[
    {
        "image_filename": "designv11_64_0003395_icarcv.2016.7838849-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003395_icarcv.2016.7838849-Figure1-1.png",
        "caption": "Figure 1. Sketch of one-dimensional reaction wheel",
        "texts": [
            " This paper is organized to introduce our proposed stabilization method (Section III) using the mathematical modeling of the reaction wheel system, the simulation environment and experimental setup as given in Section II. The proposed stabilization method implemented on our hardware platform where experimental results are given and discussed in Section IV. As a conclusion, in a Section V the results are examined. II. EXPERIMENTAL SETUP OF ONE DIMENSIONAL REACTION WHEEL Reaction wheels provide the physical means to rotate a spacecraft, based on the principle of angular momentum transfer and Newton\u2019s Third Law of action reaction. Fig.1 shows one-dimensional prototype of reaction wheel schematic design that indicates rotating axes of platform and wheel, parameters and constants given also in Table 2. A reaction wheel is not only driven by electromagnetic torque produced by brushless DC Motor, but also resisted by friction torque between the wheel and bearing, spring torque caused by the cable which moves with platform as shown on Fig.2 and disturbance torques on the space environment that are magnetic residual, solar radiation and gravity gradient",
            " In our experimental setup, the nut on Maxon DCX Brushed Motor spindle can move 37 mm in one second duration. In our case, 3 mm linear movement of nut is enough to stop the wheel. So, breaking mechanism in this system stops the wheel in 0.08 second. In this manner, using this approximately double-quick braking mechanism, the response time of the system could be decreased. If more agile braking is required, ball screw pitch can be selected higher. Although this experimental setup and product list as shown in Fig.1, Fig.2 and Table 1 is configured for one axis reaction wheel, it is also possible to customize it in order to get multiaxis reaction wheel systems. The main constants and parameters of experimental setup are listed in Table II. Fig.2 illustrates the one dimensional stabilization prototype that is built in order to assess the feasibility and develop the control algorithms for one dimensional indirect stabilization problem. The one-dimensional indirect stabilization system on which the experiments are conducted is illustrated in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002316_0959651815617883-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002316_0959651815617883-Figure16-1.png",
        "caption": "Figure 16. Results of the experimental test of the motion planning on the build prototype of a four-link planar serial robot with one moving obstacle in the workspace with the shortest path as the objective function.",
        "texts": [],
        "surrounding_texts": [
            "In this paper, a novel optimal collision-free motion planning for planar redundant serial robots was proposed, which was based on convex optimization and receding horizon concept in a disjunctive programming structure. It should be noted that the suggested algorithm benefits from a moving horizon which lead to the global optimum solution regarding the considered horizon and reduced the processing time and, thus, resulted at Middle East Technical Univ on December 31, 2015pii.sagepub.comDownloaded from in an appropriate method for real-time control of robots. Several case studies were solved based on the proposed approach for a three-link and a four-link planar redundant serial robots with two widely used objectives, i.e. minimum transition time and shortest path. In addition, it was experimentally implemented on a four-link planar serial robot. The computational time at each step was reported as less than 0.3 s. In addition, due to the predictive nature of this algorithm, the end-effector passed around the obstacles very smoothly and since at each step the problem was solved for a limited horizon, new data and positions of the dynamic obstacles were considered as well. Results demonstrated the reliability, efficacy and high computational speed of this combination which could be used for real-time purposes, compared with the previous methods reported in the literature. However, the number of horizon to be calculated is a delicate task, for which, in critical conditions, by reducing this number, more rapid decisions can be made which increases the safety of the robot. Moreover, in a safe environment, this number can be increased in order to reduce the computational time. As ongoing works, the algorithm is being extended for spatial redundant serial robots and to solve, in an online manner, the singularity avoidance problem as an additional constraint in the optimization problem."
        ]
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.14-1.png",
        "caption": "FIGURE 8.14",
        "texts": [
            " They are GSTIFF, SI2_GSTIFF, and WSTIFF. GSTIFF is the default integrator and is fast and accurate for displacements. It is used for a wide range of motion simulations. SI2_GSTIFF provides better accuracy of velocities and accelerations, but can be significantly slower. WSTIFF provides better accuracy for special problems, such as discontinuous forces. Static analysis is used to find the rest position (equilibrium condition) of a mechanism in which none of the bodies are moving. A simple example of static analysis is illustrated in Figure 8.14(a), in which an equilibrium position of a block is to be determined according to its own mass m, the two spring constants k1 and k2, and the gravity g. Very often, a static analysis is carried out to find the initial equilibrium configuration of the system before a kinematic or dynamic analysis is conducted. As discussed earlier, kinematics is the study of motion without regard to the forces that cause it. A mechanism can be driven by a motion driver for a kinematic analysis, where the position, velocity, and acceleration of each link of the mechanism can be analyzed for a given period. Figure 8.14(b) shows a servomotor driving a mechanism at a constant angular velocity. Dynamic analysis is employed for studying the mechanism motion in response to loads, as illustrated in Figure 8.14(c). This is the most complicated and common, and usually more time-consuming, analysis. In motion analysis software, the results of the analysis can be realized using animations, graphs, reports, and queries. Animations show the configuration of the mechanism in consecutive timeframes. They give a global view of the mechanism\u2019s behavior, for example, a single-piston engine shown in Figure 8.15(a). The animation may be exported to AVI for other needs. A joint or a part may be chosen to generate result graphs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002545_s11837-016-1963-5-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002545_s11837-016-1963-5-Figure2-1.png",
        "caption": "Fig. 2. Special sample holder and moving device for polishing the sample (Access e.V.).",
        "texts": [
            " To prepare the sample, commercially available equipment is appropriately modified to allow for a reproducible and well-defined ablation depth that detects the same field of interest in all serial sections. Polishing the sample is done with a polishing and grinding machine (type Saphir from ATM GmbH) equipped with a special mounting device for the sample made by Access e.V. The sample is mounted permanently in a sample holder, which is then adapted to the moving device, consisting of a height-adjuster, an additional weight, and a parallel guide (Fig. 2). For polishing, the sample holder device is adapted to the polishing machine (Fig. 3). Contact pressure was balanced by the additional weight. During polishing, the sample holder also rotates relative to the rotating grinding wheel. This allows uniform abrasion across the whole cross section of the sample. To achieve best quality preparation, the following parameters were selected: in a first step, polishing with a 9-lm diamond suspension on TexMet P (from Buehler; ITW Company) and, in a second step, polishing with Mastermet2 on PT-Chem (Cloeren Technology GmbH) cloth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002067_978-3-319-22876-1_22-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002067_978-3-319-22876-1_22-Figure4-1.png",
        "caption": "Fig. 4. The cubic spiral method for G2 continuous curve. (a) Cubic Bezier curve that can represented by angles and distances between the control points. (b) Final solution of G2 continuous method which proposed by [7] and solved by [8].",
        "texts": [
            " This paper adopts the method that was proposed by [7], which designs the cubic Bezier spiral to generate G2 continuous path by applying interpolation in the adjacent piecewise linear segments. The method introduces a cubic Bezier curve which is controlled by four control points. The cubic spline still obeys the form of equation (3) with n = 3, and let t0 = 0 and t1 = 1, then the initial and final second order derivatives are, P (2) [0,1](0) = 6(p1 \u2212 2p2 + p3), P (2) [0,1](1) = 6(p0 \u2212 2p1 + p2) (13) In [7], the proposed Beizer spiral algorithm is assumed to be composed by translation and rotation (see in Fig. 4(a)). B0, B1, B2, B3 are the control points of the cubic curve, \u03b1 is the angle between (B1 \u2212 B0) and (B2 \u2212 B1), \u03b2 is the angle between (B2 \u2212 B1) and (B3 \u2212 B2), g = ||B1 \u2212 B0||, h = ||B2 \u2212 B1||, k = ||B3 \u2212 B2||. If let B0 = (0, 0), the Bezier curve can be represented as, C(t) = (x(t), y(t)) (14) where, x(t) = 3g(1 \u2212 t2)t + 3(g + hcos\u03b1)(1 \u2212 t)t2 + (h + hcos\u03b1 + kcos(\u03b1 + \u03b2)) y(t) = 3hsin\u03b1(1 \u2212 t)t2 + (hsin\u03b1 + ksin(\u03b1 + \u03b2))t3 (15) In order to ensure curvature continuity at the joint of two curves, the parameters of the cubic Bezier can be adjusted. Analytical solution was given for this method in [8] which is illustrated in Fig. 4(b), where E1, E2, E3 are original control points for the local segment. IL1, IL2, IL3, IL4 and IR1, IR2, IR3, IR4 are the designed control points of the proposed method, where IL1 = E2 \u2212 d \u00b7 U1, IL2 = IL1 + gL \u00b7 U1, IL3 = IL2 + hL \u00b7 U1 IL4 = IL3 + kL \u00b7 UM , IR4 = E2 + d \u00b7 U2, IR3 = IR4 \u2212 gR \u00b7 U2 IR2 = IR3 \u2212 hg \u00b7 U2, IR1 = IR2 \u2212 kg \u00b7 UM (16) Where, d = 1.1228 \u00b7 sin\u03b2/Kmax(cos\u03b2)2, hL = hg = 0.346d gL = gR = 0.58hL, kL = kR = 1.31hlcos\u03b2 U1 = (E2 \u2212 E1) ||E2 \u2212 E1|| , U2 = (E3 \u2212 E2) ||E3 \u2212 E2|| , UM = (IR2 \u2212 IL3)/||IR2 \u2212 IL3|| (17) The angle \u03b2 = \u03b3/2, Kmax is the maximum curvature that allowed for the current robot"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure6.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure6.1-1.png",
        "caption": "Fig. 6.1 General configuration for a classical plate problem",
        "texts": [
            " Based on the three basic equations of continuum mechanics, i.e. the kinematics relationship, the constitutive law, and the equilibrium equation, the partial differential equation, which describes the physical problem, is derived. The weighted residual method is then used to derive the principal finite element equation for classical plate elements. The chapter exemplarily treats a four-node bilinear quadrilateral (quad 4) bending element. 6.1 Introduction A classical plate is defined as a thin structural member, as schematically shown in Fig. 6.1, with a much smaller thickness h than the planar dimensions. It can be seen as a two-dimensional extension or generalization of the Euler\u2013Bernoulli beam. The following derivations are restricted to some simplifications: \u2022 the thickness h is constant and much smaller than the planer dimensions a and b: h a and h b < 0.1, \u2022 the thickness h is constant (\u2192 \u03b5z = 0) and the undeformed plate shape is planar, \u2022 the displacement uz(x, y) is small compared to the thickness dimension h: uz < 0.2h, \u2022 the material is isotropic, homogenous and linear-elastic according toHooke\u2019s law for a plane stress state (\u03c3z = \u03c4xz = \u03c4yz = 0), \u2022 Bernoulli\u2019s hypothesis is valid, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure13-1.png",
        "caption": "Figure 13 Specimen configuration of CBT1 and CBT2 for combined bending and torsion test (see online version for colours)",
        "texts": [
            " The combined bending and torsion test is simple but important mechanical test that measures the specimen\u2019s capacity in resisting to the interaction effects of torsion in combination with bending. The FEM ABS specimen was tested in the combined bending and torsion test, and FEA study was also performed for a better understanding of the mechanical performance of the material. The specimens tested in combined bending and torsion test were fabricated in two different configurations, one is in the xz plane (CBT1) and the other is in the xy plane (CBT2) as shown in Figure 13. The specimen dimension was shown in Figure 14(a). A notch was made in the specimens in order to place the weight hanger. The tests were conducted on a TERCO Twist and Bend testing machine and the load was applied by adding weights on the weight hanger as shown in Figure 15. One side of the specimen was rigidly clamped to restrict all its degree of freedom. The dial gauge can measure as precise as 0.01 mm. For each specimen configuration, six pieces were tested, and the average deflections at 100 g weights were measured as shown in Table 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003048_ebccsp.2016.7605267-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003048_ebccsp.2016.7605267-Figure1-1.png",
        "caption": "Fig. 1. Body-fixed and inertial reference frame",
        "texts": [
            " For undirected graphs, L is symmetric and positive semi-definite, i.e., L = LT \u2265 0. The row sums of L are zero. Thus, the vector of ones 1 is an eigenvector corresponding to eigenvalue \u03bbi(G) = 0, i.e., L1 = 0. For connected graphs, L has exactly one zero eigenvalue, and the eigenvalues can be listed in increasing order 0 = \u03bb1(G) < \u03bb2(G) \u2264 ... \u2264 \u03bbN (G). The second eigenvalue \u03bb2(G) is call the algebraic connectivity. B. VTOL-UAVs model Firstly, assume that a VTOL-UAV can be modeled as a rigid body (see Fig. 1). Then, consider two orthogonal right-handed coordinate frames: the body coordinate frame, Eb = [ e b 1 , e b 2 , e b 3 ], located at the center of mass of the rigid body and the inertial coordinate frame, Ef = [ e f 1 , e f 2 , e f 3 ], located at some point in the space. The rotation of the body frame Eb with respect to the fixed frame Ef is represented by the attitude matrix R \u2208 SO(3) = {R \u2208 R 3\u00d73 : RTR = I, detR = 1}. The cross product between two vectors \u03be, \u03c7 \u2208 R 3 is represented by a matrix multiplication [\u03be\u00d7]\u03c7 = \u03be \u00d7 \u03c7, where [\u03be\u00d7] is the well known skew-symmetric matrix"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.16-1.png",
        "caption": "Figure 9.16 MTPA pu current trajectory, maximum current circles (is = 1 and is= 1.5), torque hyperbola, and voltage ellipses (with centre x) at different speeds (\u03c9s1= 1, \u03c9s2= 2, \u03c9s3= 3). The voltage ellipses are drawn taking approximately 15 % voltage reserve into account. The machine pu parameters are Ld= 0.4, Lq= 0.67, and \u03c8PM= 0.62. The current vector is is 0.9 pu at 110\u00b0 producing 0.59 pu PM torque and 0.06 pu reluctance torque with the stator flux-linkage vector \u03c8s= 0.77 at 45\u00b0. The dark gray area is the normal operating area of the drive, and the light gray area is the temporary operating area with 50 % increased current. With the rated current, stator flux linkage is reduced by 0.4 pu for a remaining 0.62 0.4= 0.22, which results in a maximum speed that is 4.5 times the rated speed at no load in field weakening. Greater torque locates the corresponding hyperbola further from the origin, while a higher speed produces a smaller voltage ellipse.",
        "texts": [
            " Field weakening control At higher speeds, the MTPA control strategy is no longer appropriate, and field-weakening (FW) control should be considered. The d-axis current must be controlled so it weakens the stator flux linkage. Per-unit current components can be calculated using a maximum voltage usmax constraint (Haque et al., 2003). The field-weakening current components are as follows. 2 L2 L2 \u03c82 L2i2 u2 =\u03c92Ld\u03c8PM Ld\u03c8PM q d PM q sFW smax s idFW (9.65) L2 L2 q d iqFW i2 i2 (9.66)sFW dFW Current component values can be calculated from these equations at different speeds and current levels isFW. Figure 9.16 demonstrates the application of both the MTPA and the FW control strategies. Figure 9.16 illustrates how torque must be reduced as speed increases. An increasing share of the electric current resources must be used to demagnetize the stator flux linkage. The centre of the voltage ellipses is located by the characteristic current ix on the negative d-axis. In this machine, ix= 0.62/0.4= 1.55. When the voltage limit is reached, the control method must be changed from MTPA to FW control. In field weakening, torque drops automatically with increasing electric current angle \u03b3 despite maintaining stator current at its rated value"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000469_978-3-319-21118-3_19-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000469_978-3-319-21118-3_19-Figure1-1.png",
        "caption": "Fig. 1 The main coordinate systems used in the quadrotor dynamic model",
        "texts": [
            " The main assumptions for dynamic model development were: \u2022 quadrotor is axially symmetric, \u2022 quadrotor is controlled in \u201cx\u201d configuration, \u2022 rotation axes of rotors are parallel to the body vertical axis of symmetry, \u2022 all parts of the quadrotor are rigid, \u2022 mass of the vehicle is constant, \u2022 induced velocity is modeled with the Glauert formula for forward flight, \u2022 equations of motion describe the motion of the body center of mass. The main coordinate systems used in the model development are shown in Fig. 1. Using assumptions above, nonlinear equations of quadrotor motion in 0bxbybzb coordinate system were formulated in a general form [20, 21]: Ip _x \u00bc f I x\u00f0 \u00de \u00fe fGK y\u00f0 \u00de \u00fe fAK x; y\u00f0 \u00de \u00fe fAW\u00f0x; di\u00de; \u00f01\u00de _y \u00bc Tx \u00f02\u00de where x \u00bc U V W P Q R\u00bd T\u2014state vector composed of linear velocities and angular rates, y \u00bc xg yg zg U H W\u00bd T\u2014vector describing position coordinates and angles of attitude, \u03b4i\u2014ith engine control signal, Ip\u2014inertia matrix, f I x\u00f0 \u00de\u2014inertia loads, fGK y\u00f0 \u00de\u2014gravity loads, fAK x; y\u00f0 \u00de\u2014fuselage aerodynamic loads, fAW x;d\u00f0 \u00de\u2014rotors aerodynamic loads, T\u2014transformation matrix in kinematic equations [20, 21]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.24-1.png",
        "caption": "FIGURE 6.24",
        "texts": [
            " As the vehicle rolls through an angle 4 the springs on each side are deformed with a displacement, ds, given by \u03b4s = \u03c6 Ls/2 \u00f06:4\u00de The forces generated in the springs Fs produce an equivalent roll moment Ms given by Ms = Fs Ls = ks \u03b4s Ls = ks \u03c6 Ls 2/ 2 \u00f06:5\u00de Calculation of roll stiffness due to road springs. The roll stiffness contribution due to the road springs KTs at the end of the vehicle under consideration is given by KTs = Ms/\u03c6 = ks Ls 2/ 2 \u00f06:6\u00de In a similar manner the contribution to the roll stiffness at one end of the vehicle due to an anti-roll bar can be determined as follows (Figure 6.24). In this case if the ends of the anti-roll bar are separated by a distance Lr and the vehicle rolls through an angle 4, the relative deflection of one end of the anti-roll bar to the other, dr, is given by \u03b4r = a \u03b8 = \u03c6 Lr \u00f06:7\u00de The angle of twist in the roll bar is given by GJ TL\u03b8 r= \u00f06:8\u00de Calculation of roll stiffness due to the anti-roll bar. where, as discussed earlier, G is the shear modulus of the anti-roll bar material, J is the polar second moment of area and T is the torque acting about the transverse section of the anti-roll bar"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.12-1.png",
        "caption": "Fig. 10.12 Effect of air gap on the magnetic circuit \u03a6 vs. ni characteristics. The air gap increases the current Isat at the onset of core saturation",
        "texts": [
            " With no air gap (Rg = 0), the inductance is directly proportional to the core permeability \u03bc. This quantity is dependent on temperature and operating point, and is difficult to control. Hence, it may be difficult to construct an inductor having a well-controlled value of L. Addition of an air gap having a reluctance Rg greater than Rc causes the value of L in Eq. (10.31) to be insensitive to variations in \u03bc. Addition of an air gap also allows the inductor to operate at higher values of winding current i(t) without saturation. The total flux \u03a6 is plotted vs. the winding MMF ni in Fig. 10.12. Since \u03a6 is proportional to B, and when the core is not saturated ni is proportional to the magnetic field strength H in the core, Fig. 10.12 has the same shape as the core B\u2013H characteristic. When the core is not saturated, \u03a6 is related to ni according to the linear relationship of Eq. (10.28). When the core saturates, \u03a6 is equal to \u03a6sat = BsatAc (10.32) The winding current Isat at the onset of saturation is found by substitution of Eq. (10.32) into (10.28): Isat = BsatAc n (Rc +Rg) (10.33) The \u03a6-ni characteristics are plotted in Fig. 10.12 for two cases: (a) air gap present, and (b) no air gap (Rg = 0). It can be seen that Isat is increased by addition of an air gap. Thus, the air gap allows increase of the saturation current, at the expense of decreased inductance. Consider next the two-winding transformer of Fig. 10.13. The core has cross-sectional area Ac, mean magnetic path length m, and permeability \u03bc. An equivalent magnetic circuit is given in Fig. 10.14. The core reluctance is Fig. 10.13 A two-winding transformer Core n1 turns + v1(t) i1(t) + v2(t) i2(t) n2 turns Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000430_978-3-319-15684-2_4-Figure4.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000430_978-3-319-15684-2_4-Figure4.5-1.png",
        "caption": "Fig. 4.5 The finite rotation angles or the robot body during walking in the down direction",
        "texts": [],
        "surrounding_texts": [
            "We describe an example of the turn angle ay to the left (see Fig. 4.6). We define longitudinal zi(T) and cross- xi(T) the coordinates of the center of gravity of the robot body at the time of the end of the current step T in the new rotated (i + 1)th coordinate system. The transition matrix to the (i + 1)th system of the ith coordinate system can be represented as \u00f04:1\u00de xi+1 xi Therefore, in the system (i + 1) transverse and longitudinal coordinates of the center of gravity can be represented as xi\u00fe1 \u00bc xi\u00f0T\u00de cos\u00f0ay\u00de ; zi\u00fe1 \u00bc 0: \u00f04:4\u00de Similarly, we write the expression for the velocity of the center of gravity in the system (i + 1), Vi\u00fe1 \u00bc Ai\u00fe1;i Vi \u00bc cos\u00f0ay\u00de 0 sin\u00f0ay\u00de 0 1 0 sin\u00f0ay\u00de 0 cos\u00f0ay\u00de 2 64 3 75 _xi\u00f0T\u00de _yi\u00f0T\u00de _zi\u00f0T\u00de 2 64 3 75 \u00bc _xi\u00f0T\u00de cos\u00f0ay\u00de _zi\u00f0T\u00de sin\u00f0ay\u00de _yi\u00f0T\u00de _xi\u00f0T\u00de sin\u00f0ay\u00de \u00fe _zi\u00f0T\u00de cos\u00f0ay\u00de 2 64 3 75 ; \u00f04:5\u00de where Ai\u00fe1;i\u2014cosine matrix, Vi\u2014column-vector of the speeds at time T onto the iaxis. Therefore, we obtain the projection of the transverse and longitudinal velocities in the rotated angle ay system i + 1 as _xi\u00fe1 \u00bc _xi\u00f0T\u00de cos\u00f0ay\u00de _zi\u00f0T\u00de sin\u00f0ay\u00de; _zi\u00fe1 \u00bc _xi\u00f0T\u00de sin\u00f0ay\u00de \u00fe _zi\u00f0T\u00de cos\u00f0ay\u00de: \u00f04:6\u00de So, rotation control is performed according to the formulas obtained in [5], adjusted for the rotation: xl \u00bc xi\u00f0T\u00de d 2 cos\u00f0ay\u00de \u00fe 1 k \u00f0 _xi\u00f0T\u00de cos\u00f0ay\u00de _zi\u00f0T\u00de sin\u00f0ay\u00de\u00de \u00f04:7\u00de where \u03b4\u2014the coefficient of stability in the transverse direction, T\u2014the end time of the current step; k \u00bc ffiffiffiffiffiffiffiffi g=L p ; L\u2014height of center of gravity; g\u2014the acceleration of free fall. We define coordinates of center of gravity of the robot body in the ith system, and let it coincide with the global [8]. The transition matrix in the global system can be represented as \u00f04:8\u00de Column-vector coordinates of the center of gravity in the global system can be represented as Ri \u00bc Hi;i\u00fe1 Ri\u00fe1 \u00bc xi\u00fe1 cos\u00f0ay\u00de \u00fe zi\u00fe1 sin\u00f0ay\u00de yi\u00fe1 xi\u00fe1 sin\u00f0ay\u00de \u00fe zi\u00fe1 cos\u00f0ay\u00de \u00fe zi\u00f0T\u00de \u00fe xi tan\u00f0ay\u00de 1 2 664 3 775: \u00f04:9\u00de Therefore xi \u00bc xi\u00fe1 cos\u00f0ay\u00de \u00fe zi\u00fe1\u00f0T\u00de sin\u00f0ay\u00de; zi \u00bc xi\u00fe1 sin\u00f0ay\u00de \u00fe zi\u00fe1 cos\u00f0ay\u00de \u00fe zi\u00f0T\u00de \u00fe xi tan\u00f0ay\u00de: \u00f04:10\u00de Figures 4.7, 4.8 and 4.9 present the results of the numerical solutions of the system of equations presented in [4], under the given conditions and parameters of the walking robot during left rotation on the second step by the angle ay \u00bc 0:3. t, c m z(t) x(t) Fig. 4.7 The longitudinal and transverse coordinates during left rotation on the second step in the global coordinate system"
        ]
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure3-1.png",
        "caption": "Fig 3 Analytical modeling of surface permanent magnet machine",
        "texts": [
            "The slip relative to rotor is 1 3 1 1 es p       ( ) , so the flux density relative to rotor is 3 3 _ 1 0 1 1 3 3 1 1 3 1 ( , ) {cos[( 1) ( ) ] 2 1 cos[( 1) ( ) ] 1 e rotor s e s e s s b t F p t s s p t s                       (15) 4) Harmonic component of armature magnetic field 4 1 0 0 1 1 ( , ) cos( ) cos( ) 1 {cos[( 1) ( ) ] 2 cos[( 1) ( ) ] e s s e s e s e s b t F p t t F p t p t                                           (16) The slip relative to rotor is 1 4 1 1 es p        ( ) , so the flux density relative to rotor is 4 4_ 0 1 4 4 1 4 1 ( , ) {cos[( 1) + ( ) ] 2 1 cos[( 1) + ( ) ]} 1 e rotor s e s e s s b t F p t s s p t s                            (17) The combination of equations (11), (13), (15) and (17) leads to the flux density relative to rotor for a machine with an eccentric rotor. _ 1_ 2_ 3_ 4_ e rotor e rotor e rotor e rotor e rotor b t b t b t b t b t          ( , ) ( , ) ( , ) ( , ) ( , ) (18) Calculate the additional eddy current losses in permanent magnet which is caused by eccentric rotor. The analytical modeling showed in Fig3 is based on the following assumptions: a) Neglect end effect; b) Neglect eddy current reaction; c) The magnetic field in air gap only has radial component. Based on polar coordinate 2-D model, the magnetic vector potential in permanent magnet can be given as _e rotorA r Bd r b t d     ( , ) (19) The eddy current in the magnet can be calculated by A J t      (20) Where is the conductivity of magnet. The combination of equations (19) and (20) leads to the eddy current density in magnets. 1 1_ 2 2_ 3 3_ 4 4_ e rotor e rotor e rotor e rotor J r s b t s b t s b t s b t           [ ( , ) ( , ) ( , ) ( , )] (21) The magnet loss caused by eccentric rotor can be calculated by 2 2 2 3 3 0 2 2 2 2 2 2 2 2 1 1 2 3 1 4 1 1 24 e eddy b a r r s s p L J J dS L R R s F s F s F s F                      * _ ( ) ( ) (22) Where L is the active length of magnet"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003119_icelmach.2016.7732831-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003119_icelmach.2016.7732831-Figure2-1.png",
        "caption": "Figure 2: Siemens NX draft simulation of pitting in the bearing\u2019s outer race",
        "texts": [
            " The resulting system responses serve as set points for the AMB. The use of impulse functions is necessary to calculate the rotor movements for inner race bearing faults accurately. Additionally, the calculation-time of obtaining the rotor movements by using the spring damper system is much less than executing FEM for every fault. To obtain a good estimation of the rotor movements in relation to the stator when an outer race fault occurs, a bearing ball is modeled in FEM which rolls over a pit of 1mm long and 100\u03bcm deep (Figure 2). The depth of the pit is a realistic starting point. The increase in severity will be a change in length, not in depth [5], [10]. Simulating the entire bearing system (outer race, inner race, cage, balls, rotor, stator. . . ) is currently not possible due to the limitations of the used FEMsoftware package. Simplifications of the system where carried out. The outer race and the inner race are represented by two solid profiles which are able to translate with respect to each other. Between those profiles, only one bearing ball is drawn which will move due to an implied translation and rotation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003152_arso.2016.7736267-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003152_arso.2016.7736267-Figure1-1.png",
        "caption": "Figure 1 redundant robot",
        "texts": [
            " Yu Chen is with the School of Mechanical Engineering and Automation, Beihang University, No. 37, Xueyuan Road, Haidian District, Beijing, China (e-mail: cy10071095@ buaa.edu.cn). Jiaxin Guo is with the School of Mechanical Engineering and Automation, Beihang University, No. 37, Xueyuan Road, Haidian District, Beijing, China (e-mail: guojiaxin@ buaa.edu.cn). 978-1-5090-4079-7/16/$31.00 \u00a92016 IEEE 116 simulation and motion simulation are used to verify the correctness of the algorithm. We sum up the main points of this paper in section V. The redundant manipulator is shown in Figure 1. The robot has seven rotating joints. According to the Denavit-Hartenberg (D-H) method, the coordinate system can be established on the joint axis, shown in figure 2. The DH parameters of the manipulator are shown in table 1. The transformation matrix can be built according to the DH parameters in Table 1.We can get the final transformation matrix from base to the end-effector. \ud835\udc477 0 = [ \ud835\udc5b\ud835\udc65 \ud835\udc5c\ud835\udc65 \ud835\udc4e\ud835\udc65 \ud835\udc5d\ud835\udc65 \ud835\udc5b\ud835\udc66 \ud835\udc5c\ud835\udc66 \ud835\udc4e\ud835\udc66 \ud835\udc5d\ud835\udc66 \ud835\udc5b\ud835\udc67 \ud835\udc5c\ud835\udc67 \ud835\udc4e\ud835\udc67 \ud835\udc5d\ud835\udc67 0 0 0 1 ] = \ud835\udc471 0 \ud835\udc472 1 \ud835\udc473 2 \ud835\udc474 3 \ud835\udc475 4 \ud835\udc476 5 \ud835\udc477 6 (4) B. Inverse kinematics Three joint axes on the shoulder meet in a point S and three joint axes on the wrist meet in the other point W"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001095_icra.2014.6907630-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001095_icra.2014.6907630-Figure1-1.png",
        "caption": "Fig. 1. Illustration of the two-dimensional quadrocopter model considered in this paper. The red arrows denote the propeller forces fp, and the blue arrows show state vector components (position p = (px, pz) and pitch angle \u03b8). For clarification, a three-dimensional view is shown on the top right.",
        "texts": [
            " To reduce the computational complexity, we thus restrict the learning to two dimensions and assume that the systematic out-of-plane disturbances are small. While in principle the approach presented herein is also suitable for three-dimensional learning, it is beyond the scope of this paper to verify this experimentally. In the following, we reduce our model to the xz-plane. In doing so, the vehicle\u2019s position is described by px and pz , and the attitude can be described by a single parameter \u03b8, which is the pitch angle. Fig. 1 shows an illustration of the two-dimensional model. For the two-dimensional case, the translational dynamics (2) reduce to p\u0308x = ax cos \u03b8 + az sin \u03b8, (9) p\u0308z = \u2212ax sin \u03b8 + az cos \u03b8 \u2212 g. (10) The evolution of the attitude is simply \u03b8\u0307 = \u03c9y, (11) and the angular acceleration (4) for the two-dimensional model is given by \u03c9\u0307y = d(f3 \u2212 f1)/Iyy + \u02dc\u0307\u03c9y,model, (12) where we assume a diagonal rotational inertia matrix I with entry Iyy for the y-axis. We define the state vector of the two-dimensional system to be s := (px, pz, vx, vz, \u03b8), (13) where vx := p\u0307x and vz := p\u0307z denote the translational velocities"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002380_transjsme.15-00563-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002380_transjsme.15-00563-Figure10-1.png",
        "caption": "Fig. 10 Winching motor on helipad",
        "texts": [],
        "surrounding_texts": [
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n3. \u30c6\u30b6\u30fc\u30c9\u65b9\u5f0f\u30d8\u30ea\u30dd\u30fc\u30c8\u306e\u958b\u767a\n3\u00b71 \u30c6\u30b6\u30fc\u30c9\u65b9\u5f0f\u30d8\u30ea\u30dd\u30fc\u30c8\u306e\u63d0\u6848\n\u524d\u7ae0\u306b\u3066\u63d0\u6848\u30fb\u5b9f\u88c5\u3057\u305f\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0cUGV\u306b\u3088\u308b\u904b\u642c\u4e2d\u306f\u8155\u90e8\u3092\u6298\u308a\u305f\u305f\u307f\uff0c\u5730\u9762\u3078\u306e\u6295\u5f71\u9762\u7a4d\u3092 \u5c0f\u3055\u304f\u3059\u308b\u305f\u3081\uff0c\u5c02\u7528\u306e\u30d8\u30ea\u30dd\u30fc\u30c8\u306b\u30d4\u30f3\u30dd\u30a4\u30f3\u30c8\u3067\u7740\u9678\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\u306a\u304a\uff0c\u8155\u90e8\u3068\u306e\u63a5\u89e6\u3092\u907f\u3051\u308b\u305f\u3081\uff0c\u30d8 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2010)\u3092\u5229\u7528\u3059\u308b\u3053\u3068\u3068\u3057\u305f\uff0ekenaf\u306f\uff0c\u512a\u308c\u305f\u969c\u5bb3\u7269\u8e0f\u7834\u6027\u80fd\u3092\u6709\u3057\u3066\u304a\u308a\uff0c\u968e\u6bb5\u3084\uff0c\u304c\u308c\u304d\u306e\u3042\u308b\u74b0\u5883\u3067\u3082\u8d70",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u884c\u53ef\u80fd\u3067\u3042\u308b\uff0ekenaf\u306e\u91cd\u91cf\u306b\u5bfe\u3057\u3066\uff0c\u30d8\u30ea\u30dd\u30fc\u30c8\u30fb\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u91cd\u91cf\u306f 10%\u7a0b\u5ea6\u3067\u3042\u308a\uff0c\u5341\u5206\u306b\u8efd\u91cf\u3067\u3042\u308b\u305f \u3081\uff0c\u8d70\u884c\u6027\u80fd\u3092\u5927\u5e45\u306b\u640d\u306a\u3046\u3053\u3068\u306f\u306a\u3044\uff0e\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u8f09\u305b\u305f\u30d8\u30ea\u30dd\u30fc\u30c8\u3092 Kenaf\u306e\u4e0a\u306b\u642d\u8f09\u3057\u305f\u30ed\u30dc\u30c3 \u30c8\u30b7\u30b9\u30c6\u30e0\u3092\u56f3 9\uff08\u53f3\uff09\u306b\u793a\u3059\uff0e\u3053\u306e\u56f3\u304b\u3089\u5206\u304b\u308b\u901a\u308a\uff0c\u9ad8\u3055\u65b9\u5411\u306b\u306f\u5927\u304d\u304f\u306a\u308b\u304c\uff0c\u5c55\u958b\u3059\u308b\u3068\u5927\u578b\u306b\u306a\u308b\u30de\u30eb \u30c1\u30b3\u30d7\u30bf\u3092 Kenaf\u306e\u5e45\u306b\u53ce\u3081\u3066\u642d\u8f09\u3059\u308b\u3053\u3068\u304c\u3067\u304d\u3066\u3044\u308b\u3053\u3068\u304c\u898b\u3066\u53d6\u308c\u308b\uff0e\n3\u00b73 \u30c6\u30b6\u30fc\u30c9\u65b9\u5f0f\u306e\u96e2\u7740\u9678\u8a66\u9a13\n\u88fd\u4f5c\u3057\u305f\u30d8\u30ea\u30dd\u30fc\u30c8\u3092\u642d\u8f09\u3057\u305f kenaf\u3068\uff0c\u524d\u7ae0\u3067\u63d0\u6848\u3057\u305f\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u5229\u7528\u3057\uff0ckenaf\u306e\u8d70\u884c\u8a66\u9a13\u4e26\u3073 \u306b\uff0c\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u96e2\u7740\u9678\u8a66\u9a13\u3092\u884c\u3063\u305f\uff0e\u5b9f\u9a13\u30b7\u30ca\u30ea\u30aa\u306f\uff0c\u307e\u305a\uff0ckenaf\u306f\u9060\u9694\u64cd\u7e26\u306b\u3088\u308a\u30c9\u30a2\u3092\u901a\u904e\u3057\u90e8\u5c4b\u306b\u4fb5 \u5165\u3057\u305f\u5f8c\uff0c\u969c\u5bb3\u7269\u306b\u963b\u307e\u308c\u305f\u5730\u70b9\u3067\u505c\u6b62\u3059\u308b\uff0e\u6b21\u306b\uff0c\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306b\u3088\u308b\u63a2\u67fb\u306b\u5207\u308a\u66ff\u3048\uff0c\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u96e2\u9678",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u3092\u884c\u3044\uff0c\u63a2\u67fb\u5f8c\uff0c\u30c6\u30b6\u30fc\u30c9\u65b9\u5f0f\u30d8\u30ea\u30dd\u30fc\u30c8\u306b\u7740\u9678\u3055\u305b\u308b\uff0e\u306a\u304a\uff0c\u672c\u8a66\u9a13\u306e\u76ee\u7684\u306f\uff0c\u6a5f\u80fd\u306e\u6709\u7528\u6027\u306e\u691c\u8a3c\u3067\u3042\u308b\u305f \u3081\uff0ckenaf\uff0c\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3068\u3082\u306b\u76ee\u8996\u306b\u3088\u308b\u9060\u9694\u64cd\u4f5c\u3067\u52d5\u4f5c\u3092\u884c\u3063\u305f\uff0e\u5b9f\u9a13\u306e\u69d8\u5b50\u3092\u56f3 11\u306b\u793a\u3059\uff0e\n\u672c\u8a66\u9a13\u3067\u306f\uff0c\u30b9\u30de\u30fc\u30c8\u30e2\u30fc\u30bf\u3092\u4e00\u5b9a\u56de\u8ee2\u6570\u3068\u306a\u308b\u3088\u3046\u5236\u5fa1\u3057\uff0c\u30c6\u30b6\u30fc\u306e\u7e70\u308a\u51fa\u3057\u30fb\u5dfb\u53d6\u308a\u901f\u5ea6\u304c\u307b\u307c\u4e00\u5b9a\u306e 150 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        ]
    },
    {
        "image_filename": "designv11_64_0001383_acc.2014.6858670-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001383_acc.2014.6858670-Figure1-1.png",
        "caption": "Fig. 1: Reference coordinate frames [3].",
        "texts": [
            " To form the basis of the relative position model, we use the standard definition of the Earth-Centered Inertial (ECI) frame Fi, with z-axis towards celestial north. We also employ a standard LVLH-definition of the leader orbit reference frame Fl, with unit vectors defined as er = rl rl , e\u03b8 = eh \u00d7 er and eh = h h , where h = rl \u00d7 r\u0307l is the angular momentum vector of the orbit, and h = \u2016h\u2016. In addition, we define a follower orbit reference frame Ff with origin specified by the relative orbit position vector p = rf \u2212rl = xer +ye\u03b8 +zeh and with unit vectors aligned with the unit vectors in Fl at all times, as shown in Fig. 1. The relative position dynamics can be written as [10] mf p\u0308 + C(v\u0307)p\u0307 + N(p, rl, rf , v\u0307, v\u0308, ul) = uf + Fd (1) where ml(mf ) denotes the leader(follower) spacecraft mass, rl(rf ) denotes the distance of the leader(follower) spacecraft to the center of the earth, ul(uf ) denotes the leader(follower) spacecraft actuator force, and v is the true anomaly of the leader spacecraft. N(p, rl, rf , v\u0307, v\u0308, ul) = D(v\u0307, v\u0308, rf )p + n(rl, rf )+ mf ml ul (2) C(v\u0307) = 2mf   0 \u2212v\u0307 0 v\u0307 0 0 0 0 0   is a skew-symmetric Coriolis- like matrix, D(v\u0307, v\u0308, rf ) = mf   \u00b5 r3 f \u2212 v\u03072 \u2212v\u0308 0 v\u0308 \u00b5 r3 f \u2212 v\u03072 0 0 0 \u00b5 r3 f   can be regarded as a time-varying potential force, n(rl, rf ) is the nonlinear term defined as n(rl, rf ) = mf\u00b5   rl r3 f \u2212 1 r2 l 0 0  "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002423_9781782421955.418-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002423_9781782421955.418-Figure2-1.png",
        "caption": "Figure 2 Measuring principle of the gear integrated error curve",
        "texts": [
            " The gear integrated error curve (GIEC) reveals the deviations of every point on all meshing flanks and the gear pair integration error curve (GPIEC) is related to the gear errors on driving and driven gears and to the non-dimensional stiffness of the teeth. 2.1 The GIEC measurement Figure 1 shows the GIEC measures the deviations of all points on gear flanks, arranged along the action line of the gear according to the meshing sequence of the correspondent mesh points. The gear integrated error curve tester using a special worm as master with high efficiency, accuracy and easy operation is used in China. The GIEC measurement is firstly introduced by Huang. Figure 2 demonstrates how it works. The tester works as simply as general single-flank testing measurement but using a special master worm, which should be at least 3 quality levels better than the gear to be tested. The master worm is of a multi-start (two or three starts) interrupted-tooth type. Only tooth flanks in one start are in operation, while others in the remaining starts are all relieve-ground. The thickness of the transmission flanks in the remaining starts shown in figure 2 is reduced to guarantee a contact ratio less than one. In figure 2 the whole pushing action is accomplished by tooth flanks in the meshing process of the master worm and the gear to be tested. The teeth on the master worm and the gear to be tested are in mesh at intervals, and there is only one pair of teeth in mesh each time. Therefore, the deviation of meshing movement is only related to single tooth error. As a result, the meshing process can take place along the full tooth height and GIEC to be obtained not only shows the full transmission error curve in each flank, but also reveals the position error in each flank simultaneously. If this process as shown in figure 2 is carried out for each flank of a gear being tested in turn, the GIEC of each flank can be gained. In the meshing process of the master worm and the gear, the whole pushing action is accomplished by tooth flanks as shown in figure 2, so only half of the flanks of the gear can be tested in the first turn of the meshing as shown in figure 3-(A). In the second turn the others can be tested as shown in figure 3-(B). If they all are marked with the same base zero and are arranged according to the meshing sequence of the correspondent mesh points, a curve like that in figure 3-(C) can be easily obtained. Then, the continuous error curve formed in this way is called a GIEC. 2.2 Comparison between GIEC and TEC The error value in GIEC designates a unique point of meshing flank of the gear, and it is called the gear integration error (GIE) of the point, which denotes the collective results of all the gear individual errors on the point along the path of action",
            "6 illustrates relevant features of the geometry of an unloaded pair of meshing involute spur gears with uniform axial lead. The two gears in presence of manufacturing errors rotate around their respective axes, the driving gear having center at O1 rotates counter clockwise. The pitch cylinder and base cylinder radii are R and bR , respectively. Tooth contact is limited to the portion of the plant of contact that lies within the addendum cylinders of the two gears as indicated by the heavy line 1 2B B . The middle of figure 2 illustrates the zone of the contact where the plane of the contact is the plane of the paper. During the process of the pair meshing, the zone of contact remains fixed, whose dimensions are determined by the active portion of the path of contact 1 2B B and the active face width B of the gears. The base pitch Pb is the length measured on the base cylinders between corresponding points on successive teeth. The lines of contact move from right to left through the zone of contact. It can be imagined that the lines on the imaginary \u201cbelt\u201d are used to define the involute tooth flanks on the two meshing gears, hence, at any given instant of time, contact between meshing pairs of teeth occurs on the lines of contact in the plant of contact. Let attach to this belt a pair of rectangular coordinates x , y as shown in the middle of figure 2. The origin of this coordinate system is placed mid-face on the line of contact of a pair of teeth pair 0j = , that\u2019s to say the value of the coordinate x designates the position of the center line of the zone of contact relative to the mid-face position of the line of contact to tooth pair 0j = . The value of the coordinate y designates the axial location of the point of contact located on the line of contact. Consider a plane cut normal to the gear axes through a pair of teeth in contact. By definition of the coordinate system, all points on this plane lie at the same value of coordinate y "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001363_2769493.2769565-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001363_2769493.2769565-Figure1-1.png",
        "caption": "Figure 1: SmartWalker hardware.",
        "texts": [
            " The walker computes the user\u2019s walking speed by detecting the user\u2019s leg movements using a laser range scanner and then combines this information with the ground inclination and the state of its brakes in the controller to compute the appropriate speed for the walker. We evaluate the walker with thirteen residents of three different retirement homes. SmartWalker [9] consists of a walker equipped with sensors and actuators and software that controls the walker. The walker that consists of a normal walking frame, two hub engines, a laser range scanner, an inclinometer, and a rotatable camera (Figure 1(a)). Located at the two rear wheels (Figure 1(b)), the hub engines contain a hall effect sensor for measuring the rotational speed of the wheel. The laser range scanner at the bottom center of the walker is a low-cost scanner1 harvested from Neato XV-11 vacuum 1400 CHF/USD for the vacuum cleaner cleaner (Figure 1(c)). It scans 360\u25e6 at 1\u25e6 resolution with the speed of 250 ms per one 360\u25e6 scan. On top of the scanner is a Pewatron PEI-Z100-AL-232-1 360\u25e6 inclinometer2 that measures the pitch of the ground. For 3D sensing, the walker has a PrimeSense Carmine 1.08 sensor, mounted on a small servo motor below the handlebar. Main processing units are a BeagleBone motherboard and a tablet computer, both running Ubuntu. The BeagleBone receives sensory data and controls actuators, and the tablet processes computationally intensive algorithms and acts as a display for the user interface"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003236_detc2016-59619-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003236_detc2016-59619-Figure4-1.png",
        "caption": "FIGURE 4. DERIVATION OF THE TUMBLE STABILITY FOR HEXAPOD MANIPULATION",
        "texts": [
            " The standard stability criteria for legged robots, which evaluates the distance between the projection of the COM and the border of supporting polygon, cant provide a sufficient measure for the amount of stability when the terrain is not a horizontal plane, although it does provide a limit which indicates whether the body is stable or unstable. The energy stability margin (ESM) are proved to be the most desirable stability criterion for walking vehicles on rough terrain [15], thus we propose a Tumble Stability Margin (TSM) for the hexapod robot manipulation based on the ESM. According to the definition of ESM, a large difference between the center of gravity and one at its highest position can evaluate the stability from the energy point of view. Here, a given support state is described as in Figure 4, the contact points are denoted as Pi. When the vehicle tumbles, the COM rotates around one of the supporting line Li, which is composed by the straight line connecting to 2 contact points. The tumble around line P1,P3 and P2,P4 will never happen because the tumbling can be prevented by the opposite supporting leg, thus we do not take these two situations into consideration. Note that the supporting lines do not need to belong to a plane, and that the tumble stability margin can be applied to mobile manipulators on arbitrary uneven terrain"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002535_tciaig.2015.2462335-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002535_tciaig.2015.2462335-Figure2-1.png",
        "caption": "Fig. 2: Shot described by five parameters (a,b,\u03b8,\u03c6,v).",
        "texts": [
            " While a player with low technical skills will generally select the easiest shot, a better player will also consider the reposition in his decision for the next shot. Thus, the best shot may not be the easiest one, but rather a more difficult one that offers a better probability of success for subsequent shots. In this section, we briefly discuss finding a balance between these two parameters, and some heuristics to facilitate the transition between each re-rack of the balls after pocketing the 14th ball. For this work, we used the Fastfiz ([7]) billiards simulator as a black-box simulator. A shot is specified by five parameters as shown in Figure 2 for which the simulator returns the resulting events and final positions of the balls after executing it. The Fastfiz simulator is deterministic, which means that when executed on the same initial table state, the same shot parameters will always result in the same table state. In reality, however, imperfection in the ball collisions, the rail\u2019s elasticity and other factors may influence the outcome of a shot. In past computational tournaments, different 1943-068X (c) 2015 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure4.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure4.16-1.png",
        "caption": "Fig. 4.16 Interpolation functions for a Timoshenko beam element with a quadratic approach for the deflection and b linear approach for the rotation",
        "texts": [
            " Quadratic interpolation for the deflection means that the deflection will be evaluated at three nodes. The linear approach for the rotation means that the unknowns will be evaluated at only two nodes. Therefore, the illustrated configuration in Fig. 4.15 for this Timoshenko element results. Evaluation of the general Lagrange polynomial according to Eq. (4.126) for the deflection, meaning under consideration of three nodes, yields 4.3 Finite Element Solution 225 (a) A graphical illustration of the interpolation function is given in Fig. 4.16. One can see that the typical characteristics for interpolation functions, meaning Ni (xi ) = 1 \u2227 Ni (x j ) = 0 and \u2211 i Ni = 1 are fulfilled. With these interpolation functions the submatrices K 11, \u00b7 \u00b7 \u00b7 , K 22 in Eq. (4.129) result in the following via analytical integration as: K 11 = ksAG 3L \u23a1 \u23a3 7 \u22128 1 \u22128 16 \u22128 1 \u22128 7 \u23a4 \u23a6 , (4.140) 226 4 Timoshenko Beams K 12 = ksAG 6 \u23a1 \u23a3 \u22125 \u22121 4 \u22124 1 5 \u23a4 \u23a6 = (K 21)T, (4.141) K 22 = ksAGL 6 [ 2 1 1 2 ] + E Iy L [ 1 \u22121 \u22121 1 ] , (4.142) which can be assembled to the principal finite element equation by making use of the abbreviation \u039b = E Iy ks AGL2 : ksAG 6L \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 14 \u221216 2 \u22125L \u22121L \u221216 32 \u221216 4L \u22124L 2 \u221216 14 1L 5L \u22125L 4L 1L 2L2(1 + 3\u039b) L2(1 \u2212 6\u039b) \u22121L \u22124L 5L L2(1 \u2212 6\u039b) 2L2(1 + 3\u039b) \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 u1z u2z u3z \u03c61y \u03c63y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 F1z F2z F3z M1y M3y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000640_icra.2015.7139415-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000640_icra.2015.7139415-Figure3-1.png",
        "caption": "Figure 3. Z shaped frame",
        "texts": [
            " The propellers have to be kept in such a way that they should not interfere with their neighbors and this was achieved by keeping them in a different plane, offset by a distance s as shown in Fig. 2. 978-1-4799-6923-4/15/$31.00 \u00a92015 IEEE 1692 It is a two layered design, where the layers are vertically separated by an offset distance allowing propellers to overlap. Each layer accommodates two rotors each of diameter d, which are placed diagonally opposite to each other. Upper layer is a Z shaped frame as shown in Fig. 3 with two arms on either sides of the stem connecting them. Two motors are mounted at the terminals of each arm. Motors 3 and 4 are mounted in upper layer and motors 1 and 2 in lower layer. Each \u201eZ - frame\u201f is characterised by the parameters a and b as shown in Fig. 3. These two layers are vertically separated by a distance and they are connected using two standoffs seperated horizontally by a distance . This distance can be calcuated using (3) such that the propeller does not interfere with the standoffs. ( ) (1) ( ) (2) \u221a (3) where, t is the propeller overlap and is an extra safety distance decided based on the tolerance of propeller. As explained above, the two key parameters in the design of VOOPS are the vertical offset and blade overlap. One of the key issues in arriving at an accepatble offset is the blade deflection during propeller rotation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003407_smc.2016.7844632-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003407_smc.2016.7844632-Figure2-1.png",
        "caption": "Fig. 2. Hex Rotor body frame with forces and torques",
        "texts": [],
        "surrounding_texts": [
            "In the following, a closer look will be taken onto the three controllers, namely attitude, velocity and position control. To get a better understanding of their function, Figure 3 serves as an overview of the inputs and outputs of the respective controllers. It is a cascaded control structure, with three controllers. Every controller has a defined input, being the errors of the states they have to control. The outputs serve as inputs for the next controller and define the desired values, forming the errors together with the states, measured by the INS. The only exception is made feeding forward the throttle command, defining the overall thrust. This command is directly fed forward to the hex rotor system, skipping the attitude control. Of course, this is necessary, as the overall thrust does not have any influence on the attitude."
        ]
    },
    {
        "image_filename": "designv11_64_0003475_icieam.2016.7910960-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003475_icieam.2016.7910960-Figure2-1.png",
        "caption": "Fig. 2. Verticality deviation. \u03b1 - the angle of the link from the vertical; \u03b8 - drive shaft turning angle. There are flat hinges, which make possible turning of the link, the rod and the wheel at points \u041e, \u0410 and \u0412, respectively. L \u2013 the distance between \u041e and \u0410; l \u2013 the rod length; d \u2013 the drive shaft radius; \u03b4 - the angle part of the link \u041e\u0410 from \u03b1 ;",
        "texts": [
            " Computational solution can be founded by RepinTretyakov procedure [10, pp. 499-503]. II. STATEMENT OF THE PROBLEM In this work the method, developed in [8, 13-15] for holonomic systems with redundant coordinates is used for solving steady motion stability and stabilization problem for a mobile four-wheeled single-link manipulator (fig. 1). The task for the manipulator is keeping its clamp on a specified height, while moving rectilinearly. The link position is determined by verticality deviation (fig. 2). The link is connected with the drive shaft by a fixed length rod, so there is a nonlinear geometric constraint between turning angles of the link and the drive shaft. For simplicity, the wheels are supposed to be deformed along the vertical only. To construct an accurate mathematical model for the mechanical part of the manipulator dynamic equations in M.F. Shul\u2019gin\u2019s [5] form are used. It is possible to assume the turning angle of the link or the drive shaft as a redundant coordinate. To insulate the model we have to include the electric drive equation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003424_ecce.2016.7854806-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003424_ecce.2016.7854806-Figure6-1.png",
        "caption": "Fig. 6: Comparison of the flux density between a healthy machine and a machine with 20% of the turns in phase A conductors are shorted.",
        "texts": [
            " Based on [18], the steady state d and q axis voltages in the case of turn-to-turn short circuit fault are given by (6)\u2212 (9): Vdsh = Vd + Vdf (6) Vdf = 2 3 (Rf cos(\u03b8)if \u2212 \u03c9e(Mhf + Lf ) sin(\u03b8)if ) (7) Vqsh = Vq + Vqf (8) Vqf = 2 3 (Rf sin(\u03b8)if + \u03c9e(Mhf + Lf ) cos(\u03b8)if ) (9) where if = If sin(\u03b8+ \u03b8f ), If is the magnitude of the current in the shorted turns, \u03b8 is the rotor position, \u03b8f is the phase angle of the current in the shorted turns, Mf is the mutual inductance between the healthy and faulted coils, and Lf is the self inductance for the shorted turns. The equivalent model will contain two components: a healthy component and a variable component related to the short circuit fault. The variable component contains a DC component and an oscillated component. The DC component depends on the number of shorted turns and on the short resistance. As the severity of the fault increases, this DC component will increase, causing the value of the d and q voltages to increase too. Fig. 6 shows a comparison of the flux density and the flux lines between healthy machine, and a machine with 20% of the total turns in two of phase A conductors are shorted. In the case of turn-to-turn short circuit fault, the total flux density will decrease in the shorted conductors, which means a decrease in the value of \u03bbq . However, in the short region, more flux lines will be closed in the d axis of the machine causing \u03bbd to increase, as shown in Fig. 6. Decreasing \u03bbq and increasing \u03bbd will increase the value of Vd and Vq . So in the case of short circuit fault, the point (Vd, Vq) in the Vd-Vq plane will shift to the top right. It is noticed that turn-to-turn short circuit fault depends on the number of the shorted turns and the values of the shorted resistance. As the number of shorted turns increases or the value of the shorted resistance decreases, the values of Vd and Vq will increase more. To find the points where the machine will be operating at maximum torque, simulations and experimental tests were performed to characterize the tested motor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001572_nme.4643-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001572_nme.4643-Figure15-1.png",
        "caption": "Figure 15. Postbuckling pattern of an anisotropic infinite film on soft substrate with loading conditions: (a) N\"x D 0; N\"y D 0; N xy D 0:010, (b) N\"x D 0:03; N\"y D 0:01; N xy D 0:010, (c) N\"x D 0:01; N\"y D 0:02; N xy D 0:010, and (d) N\"x D 0:03; N\"y D 0:03; N xy D 0:010. Note that the converged representative volume elements in (b), (c), and (d) are not rectangle.",
        "texts": [
            " To illustrate this, the following anisotropic material constants are adopted in the simulation. Young\u2019s modulus E11 D 1:2 107ksh; E22 D 1:2 106ksh, shear modulus G12 D 8:4 105ksh, Poisson ratio 21 D 0:30. Copyright \u00a9 2014 John Wiley & Sons, Ltd. Int. J. Numer. Meth. Engng 2014; 98:445\u2013468 DOI: 10.1002/nme The postbuckling pattern under four loading conditions is simulated by our adaptive FEM and shown in the relationship between postbuckling patterns and compression strains. The solid lines divide the plane into three regions with different patterns. Figure 15, when a shear strain is applied without compression strains as in the relationship between postbuckling patterns and compression strains. The solid lines divide the plane into three regions with different patterns. Figure 15(a), 1D buckling pattern is obtained. The RVE remains to be a rectangle, but its orientation changes. It is noticed that the orientation angle ' (Figure 12) is 61.7\u0131 with respect to the principal axis of the material, which is not in the direction of principal strain (45\u0131). In the relationship between postbuckling patterns and compression strains. The solid lines divide the plane into three regions with different patterns. Figure 15(b)\u2013(d), the RVEs are no longer rectangles. It is a new postbuckling behavior, which cannot be simulated by FEM with fixed rectangle RVE. We have developed an adaptive periodical RVE to simulate periodical postbuckling behavior of stiff beam or film on soft substrate under longitudinal compression. By introducing the dimensions of RVE as additional freedoms, our adaptive RVE can capture the real postbuckling pattern with an arbitrary choice of initial RVE, which can evolve itself together with the nodal degrees of freedom Copyright \u00a9 2014 John Wiley & Sons, Ltd"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000528_icra.2015.7140025-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000528_icra.2015.7140025-Figure1-1.png",
        "caption": "Fig. 1 Prototype of Intelligent Cane",
        "texts": [
            " Owing to this behavior, a swing leg on the turning direction side freely moves to the turning direction, while a swing leg on the opposite direction side of turning hardly move to the turning direction. As a result, the tandem stance is avoided. The rest part of this paper introduces a mechanism and adaptive and asymmetric admittance control of the intelligent cane robot system, and experiments of tandem stance prevention. II. INTELLIGENT CANE ROBOT SYSTEM A prototype of intelligent cane shown in Fig. 1 is developed to help the elderly walking and training their walking ability. This cane robot was designed in small-size. The cane robot consists of an omni-directional mobile base, an aluminum stick, minicomputer, and sensor system including a 6-axis force/torque sensor, and a laser range finder (LRF). The omni-directional mobile base comprises three commercially available omni-wheels and actuators, which are commercially available from Soai co.,ltd. In spite of the small size, a load capacity of this mobile base is up to 100 [kg]",
            " In the high-level supervisor, the behavior of the user is monitored based on the data obtained by the above-mentioned sensors, and the intentional direction (ITD), in which the user intends to move, is estimated based on the monitored behavior of the user. Based on the estimated intentional direction, the behavior of the cane robot is determined [9]. In the low-level motion controller, the cane robot controls its position based on the determined behavior, using admittance control method. The motion of the cane robot is three degree of freedom (DOF) consisting of translational motions and a rotational motion. In the case of translational motions, the cane robot moves in the y-axis direction that is ITD, in the x-axis direction, as shown in Fig. 1. In the case of rotational motion, the cane robot turns around z-axis that is the axis perpendicular to the ground, as shown in Fig. 1. Both of the translational motions and the rotational motion are controlled by the admittance control. When the translational motions are controlled by the admittance control, a horizontal force applied to the cane robot becomes an input of the control algorithm and a velocity of the cane robot is an output of the control algorithm. By controlling the velocity, the cane robot moves in a manner based on translational movements represented by[ Mvx 0 0 Mvy ][ v\u0307x v\u0307y ] + [ Dvx 0 0 Dvy ][ vx vy ] = [ fhx fhy ] (1) where [fhx fhy] T are horizontal forces applied to the cane robot, [vx vy] T are velocities of the cane robot, [Mvx Mvy] T are virtual masses, and [Dvx Dvy] T are virtual dampers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002586_s1052618816020096-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002586_s1052618816020096-Figure1-1.png",
        "caption": "Fig. 1. Dynamic model of f lexible rotor in elastically fixed stator (when rolling).",
        "texts": [
            " As a result, in the vicinity of the first resonance the rotor with the distributed mass behaves approximately as a disk with equal mass in the middle of a light shaft with equivalent stiffness and corresponding internal friction. In the case of small angular clearance (to the stator), we can limit ourselves to consideration of the problem of rotor travel in the linear approximation and with the respect to the plane model; i.e., we can assume that total points of the system are displaced only parallel to the \u041e\u0445\u0443 plane. Hence, when studying rotor rolling over the stator, we use the model shown in Fig. 1. Let the deviation of the rotor axis in the Oxyz reference system be designated xR and \u0443R, its rotation angle through \u03c9t, stator axis deviation through \u0445S and yS, and center-of-inertia displacement xG = xR + acos\u03c9t, yG = yR + asin\u03c9t, where a is the rotor eccentricity. Here, it is assumed that the rotor is connected to significant rotating masses due to which its rotational speed can be considered constant during the rolling. In the accepted notations the kinetic energy of the precessions of the rotor and the stator is where mR, mS are masses of rotor/stator; and IG is the moment of inertia of the rotor relative to the straight line parallel to the z axis and passing through its mass center",
            " Equality (1) subject to these dependences is The dissipative function which takes into account friction between the system and the environment is By analogy with (2), the external viscous friction coefficients can be given as dR = kR\u03a8E/2\u03c0|\u03a9| and dS = kS\u03a8S/2\u03c0|\u03a9|, where \u03a8E is the coefficient of rotor vibration energy absorption by the environment, and \u03a8S is the coefficient characterizing losses in the stator material. The expressions and generalized forces constructed for the corresponding Lagrange equations, as constant G (e.g., the influence of rotor weight G = mRg), contact normal force N, and tangential force \u0422 (Fig. 1, the opposite action of N and \u0422 on the stator where is not shown, but implied), yield the following equations of the stationary contact motion of the system in Cartesian coordinates: Equations for rolling the f lexible imbalanced loaded rotor over the elastically fixed stator are reasonably to be reduced to the form: (3) where zR = xR + iyR and zS = xS + iyS are displacements of the rotor and stator axes in complex form. The possibility of rotor rolling over the stator, which represents pure rolling or rolling with rotor slippage over the stator, suggests the existence of a periodic solution (4) where \u03a9 is the retrograde precession velocity of the rotor axis, which is equal either to the known value \u03a9rol = \u03c9r/\u03b4 or some value Qslip, which is determined by the real and dynamic friction coefficients and are further designated \u03bc and \u03bc0 = T/N, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001873_amm.813-814.1007-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001873_amm.813-814.1007-Figure2-1.png",
        "caption": "Figure 2: Thrust runner and bearing Figure 3: Different shapes of foils",
        "texts": [
            " Diameter of bearing 60mm Diameter of thrust runner 60mm Bearing and runner Material Aluminum alloy Foil thickness 0.1mm Foil material Uncoated Copper Pad angular extent 90 0 Drive unit Thrust runner Foil Bearing S-type load cell Speed sensor Different axial foil thrust bearings were tested out to obtain load carrying capabilities in order to study the effect of different geometric parameters. Detailed experiments were conducted with parametric variation to the extent possible. The figure 1 shows schematic of the test rig and figure 2 shows the direction of rotation of thrust runner and bearing. As mentioned earlier, parametric experimentation in terms of numbers of foils axially and circumferentially, angle subtended by each foil, shape of foils,were conducted essentially to study their effect on the load carrying capacities, static stiffness and dynamic stiffness conditions of the axial foil thrust bearings. The results are presented graphically and are discussed below. The operating parameters varied included are speed and the overall gap between the runner and the foil bearing base"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001661_1350650114540626-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001661_1350650114540626-Figure2-1.png",
        "caption": "Figure 2. MTM ball-on-disc geometry.",
        "texts": [
            " One observation from the earlier work was the film thickness in the incipient starved regime was time dependent; it decreased with over-rolling as more lubricant was expelled from the track, and would often take a few minutes to achieve a final value. In the current study, the starvation is monitored through the friction over 300 s intervals. at University of Hawaii at Manoa Library on July 7, 2015pij.sagepub.comDownloaded from The Mini Traction machine (MTM PCS Instruments, London, UK) was used to carry out the starvation tests (Figure 2); this employs a ball-on-disc test geometry with the following characteristics (Table 1). The ball has a diameter of 19.05mm and an rms roughness of 0.06 mm, while the disc has a diameter of 46mm and an rms roughness of 0.12mm. Both components are made of 52100 steel. The test lubricant is a commercial fully formulated 15W40 mineral oil. An additive containing oil was chosen to minimize surface damage during the starved lubrication. The test pieces are driven independently by electric motors and friction is measured by force transducers mounted on the ball shaft"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.26-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.26-1.png",
        "caption": "Fig. 4.26 a Two bodies in the state of separation; b two bodies in the state of contact (indentation, \u03b4)",
        "texts": [
            " The vector l that connects the points Pi and Pj can be evaluated by l \u00bc ri \u00feAis0Pi rj Ajs0Pj \u00f04:96\u00de The magnitude of this vector is l \u00bc ffiffiffiffiffi lT l p \u00f04:97\u00de The unit vector along the spring-damper-actuator element is defined as u \u00bc l l \u00f04:98\u00de The time rate of change of the damper length can be obtained by differentiating Eq. (4.97), yielding _l \u00bc lT _l l \u00f04:99\u00de where _l, in turn, is found from Eq. (4.96) _l \u00bc _ri \u00fe _/iBis0Pi _rj _/jBjs0Pj \u00f04:100\u00de in which Bk \u00bc sin/k cos/k cos/k sin/k ; \u00f0k \u00bc i; j\u00de \u00f04:101\u00de Figure 4.26a shows two bodies with convex profiles of a generic multibody system in the state of separation. The vector that connects the two potential contact points, Pi and Pj, is a gap function that can be written as d \u00bc rPj rPi \u00f04:102\u00de where both rPi and rPj are described in global coordinates with respect to the inertial reference frame, that is rPk \u00bc rk \u00feAis0Pi \u00f0k \u00bc i; j\u00de \u00f04:103\u00de in which ri and rj represent the global position vectors of bodies i and j, while s0Pi and s0Pj are the local components of the contact points with respect to local coordinate systems. The planar rotational transformation matrices Ak are given by Nikravesh (1988) Ak \u00bc cos/k sin/k sin/k cos/k \u00f0k \u00bc i; j\u00de \u00f04:104\u00de In turn, Fig. 4.26b depicts the same two bodies in the state of contact. As it was presented in Chap. 2, the pseudo-penetration or indentation, \u03b4, can be evaluated from the state variables of the multibody system as d \u00bc ffiffiffiffiffiffiffiffi dTd p \u00f04:105\u00de The velocities of the contact points can be expressed as _rPk \u00bc _rk \u00fe _Aks0Pk \u00f0k \u00bc i; j\u00de \u00f04:106\u00de in which the dot denotes the derivative with respect to time. The relative velocity between the contact points is projected onto the tangential line to the colliding surfaces and onto the normal to colliding surfaces, yielding a relative tangential velocity, vT, and a relative normal velocity, vN"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001604_aieepas.1958.4499867-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001604_aieepas.1958.4499867-Figure13-1.png",
        "caption": "Fig. 13. Doublesuspension bundled asembly with circular yoke-plete shield",
        "texts": [
            " To fulfill the requirements of shielding the clamp, yoke plate, and other connecting hardware, as well as providing grading for bottom units of insulator strings, a corona shield, shown in Fig. 10, was designed and tested. The RIV curves, Figs. 11 and 12, demonstrate the effectiveness of this type of shield. Shielding of Double-SuspensionString Bundled Arrangements Frequently, transmission-line designs require double strings of insulators in suspension at specific locations. A typical assembly showing a yoke-plate corona shield is shown in Fig. 13. The circular shield, open at one end, was tested on this assembly in the upright and inverted positions. These results are shown in Fig. 14. The corona shield, Fig. 10, was installed on a double-suspensionstring assembly. The corona and RIV results are illustrated by the curves in Fig. 15. Discussion of Data A comparison of the corona and RIV on single and bundled conductors is shown by Fig. 2. The bundled conductors of both diameters (1.05 inch and 1.315 inch) show a substantial decrease in noise level over the 1",
            "05-inch-OD conductor while no corona was observed on the shielded assembly using 1.90-inchOD conductors. The reduction in RIV noted for the 1.9-inch-OD unshielded conductor assembly when compared with the unshielded 1.05-inch-OD assembly was approximately 39% at critical voltage. The reduction in noise level between the foregoing respective conductors when used O-Unshielded assembly +-Shieided dssembly 92Kaminski, Jr.-Corona Shields for Suspension Assemblies 92 APRIL 1958 as bundled assemblies is 22% at critical voltage. The double-suspension-string assembly, Fig. 13, when tested with yoke-plate shield in two different positions, Figs. 13(B) and (C), does not yield as much improvement as can be realized when using the corona shield shown in Fig. 10. Referring to the curves shown in Fig. 14, a reduction in noise level of 53%, and the complete absence of corona is achieved at critical voltage with the special corona shield. With reference to the curves in Fig. 15, an unshielded bundled-conductor assembly, using double-suspension-string assemblies, is slightly better from an RIV standpoint than are single-suspension-string assemblies",
            " A reduction in RIV level of at least 50% is realized in shielded bundled-conductor double-string-suspension assemblies when compared with similar unshielded assemblies. 4. Corona-free bundled-conductor assemblies at 300-kv line-to-ground voltage can be realized with proper shielding. 5. While laboratory tests are made under controlled conditions, and it is known that the elements wil materially affect the performance of transmission lines with respect to corona and RIV, the margin of improvement demonstrated by the use of corona O-Unshielded assembly, Fig. 13(A) * Yoke-plate shield in position, Fig. 13(B) +-Yoke-plate shield in position, Fig. 13(C) &-Special corona shield, Fig. 10 Inch-OD conductor * -Unshielded single-suspension string O-Unshielded double-suspension string G-Shielded double-suspension string +-Shielded single-suspension string shields fully justifies their use in higher voltage bundled applications. References 1. OUTPUT AND REOULATION IN LONG-DISTANCE Lums, Percy H. Thomas. AJEE Transactions, vol. 28, pt. I, 1909, pp. 615-40. 2. THREE-PHAsz MULTIPLE-CONDUCTOR CRtcurr, Edith Clarke. Ibid., vol. 51, Sept. 1932, pp. 809-23"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003907_jrproc.1953.274240-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003907_jrproc.1953.274240-Figure3-1.png",
        "caption": "Fig. 3",
        "texts": [
            " Moullin, \"The use of Bessel functions for calculating the self-inductance of single layer solenoids,\" Jour. IEE (London), vol. 96, p. 133; 1949. (3) The mutual inductance M between two turns distant z apart can be obtained from tables in footnote refer-_ 10 The case where the coil is earthed at one end can be directly obtained, and it is given in Appendix C. A pril538 1 d 2 2 + - X 10-3 IuH. 16 0 Mostafa and Gohar: Characteristics of Single-Layer Coils ence 11. Plotting Mla to a base of y/a and remember that M(y) = M(-y), Fig. 3 is obtained, where y is: distance between the two turns which require M. 1 interesting values of M/a are those starting from y= At y = 0, the ordinate is I/a. The summation in should be carried out turn-by-turn. However, unless number of turns is small, the same result can be sa factorily obtained by an integration process. Therefc (2) reduces to c b/2 @ r z-DAv =jclix+ jJ Midz + j- Mi,dz D z+D D _b/2 p b/2 = - | Mlizdz. D J-b/2 ing the Fhe *D. (2) the tis)re, where n is an odd integer. The upper limit of the summation will be given later in the paper. A convenient representation of the characteristic, Fig. 3, is a trigonometric polynomial by means of Fourier analysis with 2b/a as its fundamental period. Mi/a = Ao + E Am cos (mry/b). m=l (6) The upper limit of the summation in (6) depends on the number of terms in the expansion required for satisfactory representation. Ao and Am are to be determined from Fourier analysis. It is clear that Ao and Am depend on D/a, b/a, and d/a. (4) In order to have the same M1/a- y/a characteristic for all practical coils, a value of 0.005 for D/a and a value of 0.01 for d/a will be assumed. For other ratios, the representation of the characteristic will be practically unaffected (Appendix B). M1 is the curve 1 2 2' 3' 3 4 (Fig. 3). From the nature of the problem, i2 can be represented as i2= ao + E an cos (nrz/b) n=l \"i F. E. Terman, \"Radio Engineering Handbook,\" McGrawHill Book Co., Inc., New York, N. Y.; 1943. (5) gives the inductance mx between a turn at a distance x and all other turns of the coil, the integration of which along the coil gives L,. b2 4a/r2 Lc = - Ao + EZAn/T2 uH. D27r2 4 n-i (8) L. is calculated by (8) for several coils having different dimensions, and the results are in good agreement with those of Nagaoka (Table 1)",
            " Effect of Using Ratios for D/a and d/a other than 0.005 and 0.01, Respectively The effect of using a value for Dla other than 0.005 will be first considered. The pitch to diameter ratio in almost any practical coil is not less than 0.005, which is the value considered in the text. If the summation process is carried out turn-by-turn, the result will not therefore be affected by any other value for Dla. However, if an integration process is adopted, the result will be slightly affected. As the representation of Fig. 3 is a Fourier one, the summation can be easily carried ouit. As an example, sin D )sin-n-1 2 2 E sin mDv, m=0 sin D/2 while IJD D J lnD 2 /D sin DdD = sin D D 2 Y2 which is practically the same provided the number of turns is not very small (nD/2=7r/2). Therefore, for large number of turns, the integral in (4) and the like could be considered practically unaffected by varying the ratio D/a provided N is large. The effect of using a value for dla other than 0.01 can also be practically neglected, and can be clarified as follows: In determining Ao, Am in the text, the 48- ordinate scheme was adopted"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure13-1.png",
        "caption": "Fig. 13. Proposed Cooling Fan for Brake-Path",
        "texts": [
            " Attempts to direct jets of air on to the brake-path have not been very successful, the reason probably being that the actual weight of air emerging from the jet has been comparatively small. What is required is a fan which will pump a really substantial weight of air through the inside of the brake. The only feasible way of doing this appears to be to form a centrifugal impeller on the drum backplate, drawing air through the centre of the road wheel and discharging it through orifices near the outer periphery of the brake carrier-plate, as in Fig. 13. The effective- a End view of dnun and fan. b SectionAA. ness of such a scheme is seriously limited by the low rotational speed of the road wheel, particularly on \u2019buses and trucks, where improved cooling is most needed, but it appears to be the only feasible way by which air can be circulated within the brake. Forward facing air scoops are used on racing cars, but are ineffective on slow-moving vehicles with shrouded wheels. As far as is known, the rate of wear of all rubbing solids increases with rise of temperature, unless, of course, the coefficient of friction falls so much that the materials save themselves, preventing any further temperature rise",
            " Those problems presented little difficulty on small and medium-sized family cars, but great difficulty on high-speed, high-performance cars, owing mainly to the use of smaller and wider wheel rims which gave less air space round the drums, and to the modem styling of wings which necessitated more cowling of the brakes. The author had referred to public transport vehicles in which the heat absorbing powers of much thicker drums had given beneficial results, but that method could not be used extensively on modern high-speed cars as the unsprung mass would have adverse effects on suspension and steering. Where expense was a minor consideration, the use of bimetal cast-iron drums having cast aluminium fins would give improved cooling. He was doubtful about the method of cooling shown in Fig. 13 and he asked whether that had been tried out. Some Italian sports cars had louvred wheel disks over wire wheels to induce air flow outwards into the slipstream of the vehicle, but he believed that they were more ornamental than useful. Air scoops were liable to pick up water and sand when placed in the most advantageous position for cooling. It would appear that for the time being improvements could only be obtained from the use of well finned drums, wider shoes having lower unit pressures, and on moulded liners which had a greater resistance to fade",
            " Heat dissipation, assuming the temperature to be constant, was dependent upon the surface area, and he would have thought that ribbing the surface was an economical method of increasing the surface area. Presumably, heat dissipation could also be varied by the use of bi-metal drums, in which heat dissipation to the atmosphere could be increased, but that involved the question of heat gradient throughout the drum, and particularly at the joining of the two metals. He presumed that the cooling fan shown in Fig. 13 was a theoretical proposition rather than a practical one, except possibly for racing cars, as no provision had been made for variation in climatic conditions. In a vehicle fitted with such an arrangement it would, in his opinion, be necessary for the driver to have a regular weather forecast. He asked what effect moulding the lining completely on to the shoe, instead of riveting, would have on heat dissipation. A riveted lining mobably allowed for slight local expansion. Mr. H. F. ELLIS, A.M.I",
            " It was desirable that the characteristics of brake linings should be known accurately, and he would be interested to hear how that information should be obtained. He inquired about the best method of determining the temperature at the skin of a brake lining. The usual method of placing the thermocouple under the surface would not appear to give the true temperature on the working face immediately in contact with the brake path. He also asked whether the author considered it was best to employ a comparatively st i f\u20ac brake drum, or a relatively flexible drum which would conform to brake shoe deformation. In Fig. 13, showing the method of cooling a brake, the entry and exit slots appeared to be rather small and the path taken by the cooling air seemed rather hazardous to be fully effective. The statement that brake fade could become a stable condition was a little puzzling. A leakage of oil or warm grease from hub bearings was more important than appeared to be generally appreciated. If the difficulties in producing an effective seal were too great, designers should ensure that the brake linings and brake path were not contaminated with the leaking oil"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001226_06401.0133ecst-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001226_06401.0133ecst-Figure1-1.png",
        "caption": "Figure 1. Chemical structure of paraoxon and carabaryl",
        "texts": [
            " Carbaryl, multi-walled carbon nanotubes, lyophilized salmon sperm DNA salt, N-hydroxysulphosuccinimide (NHS), N-ethyl-N-(3dimethylaminopropyl) carbodiimide hydrochloride (EDC), polyethyleneimine (PEI), 2- (N-morpholino) ethanesulfonic acid (MES), N-cyclohexyl-2-aminoethanesulfonic acid (CHES), phosphate buffered saline (PBS), tris(hydroxymethyl)aminomethane (TRIS) were all obtained from Sigma-Aldrich. Ultrapure DI water (Millipore resistivity- 18.2 M\u2126 cm-2) was used for all the sample preparations. 2.2 Instruments All electrochemical measurements were recorded with a CHI 660 (CH Instruments, Austin, TX) potentiostat connected to a computer utilizing the chi990b software package. The flow injection analysis system for amperometric measurement is shown in Figure 1. Cyclic voltammetric (CV) measurements were performed using a three-electrode system consisting of glassy carbon electrode (GCE, 3-mm diameter), Ag/AgCl (3 M KCl) and platinum wire as a working, reference and counter electrodes respectively. Scanning Electron Microscope (SEM; JSM-7000F field emission SEM, JEOL Ltd, Japan) was used to characterize the electrostatically self-assembled interfaces on GC electrode. 135 ) unless CC License in place (see abstract).\u00a0 ecsdl.org/site/terms_use address. Redistribution subject to ECS terms of use (see 138"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure7-1.png",
        "caption": "Fig. 7 Von Mises stress on the optimized caliper.",
        "texts": [
            " Structural performances of the obtained solution are reported in Tab. 2 and compared with actual components. Tab. 2 Structural performances of the obtained solution and comparison with the actual component. Performance index Value Variation w.r.t. actual component Caliper mass [kg] 1.063 -22 % Upright mass [kg] 1.529 -23 % D1 [mm3] 375 -15 % D2 [mm3] 201 -47% D3 [mm3] 192 -42% A preliminary stress analysis on the obtained shapes has been performed for evaluating the structural integrity, the obtained stress field is depicted in Fig. 7 and Fig. 8. The obtained stress levels are quite acceptable in all the components, obviously the geometry has to be refined and more detailed evaluation is required, maybe with a kind of shape optimization in the critical areas for reducing the stress values at the notches [10, 11]. In any case the undoubtful advantage is that an optimized solution has been obtained at a very early stage of the design process. The obtained solution is a very good starting point for the development of the new components"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001270_s40435-014-0064-y-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001270_s40435-014-0064-y-Figure1-1.png",
        "caption": "Fig. 1 Scheme of weight-fibre-pulley-drive mechanical system with asymmetric position of weight",
        "texts": [],
        "surrounding_texts": [
            "Originally it was supposed that for the experimental measurement focused on determining properties of a fibre an inverted pendulum driven by two fibres attached to a frame would be used. Its properties were investigated very thoroughly applying calculation models (see e.g. [8]). But strength calculation results drew attention to a high loading of fibres which were to be used in the experimental measurements (carbon or wattled steel wire) and to the possibility of their breaking [9]. Due to those reasons a different mechanical system was chosen for the experimental measurements (its geometrical arrangement was changed several times on the basis of various pieces of knowledge). Experimental measurements focused on the investigation of a fibre behaviour are performed on an assembled weigh-fibre-pulley-drive mechanical system (see Figs. 1, 2). A carbon fibre with a silicone coating (see e.g. [9]) is driven with one drive and is led over a pul- ley. The fibre length is 1.82 m (fibre mass is 4.95 g), the pulley diameter is 80 mm. The position of the weight can be symmetric [3] or asymmetric with respect to the vertical plane of drive-pulley symmetry (in this article asymmetric position is considered). Distance of the weight from the vertical plane of drive pulley symmetry is 280 mm in the case of the asymmetric weight position (see Figs. 1, 2). At the drive the fibre is fixed on a force gauge. On the other end of the fibre there is a prism-shaped steel weight (weight 3.096 kg), which moves in prismatic linkage on an inclined plane. The angle of inclination of the inclined plane could be changed (in this case the angle is \u03b1 = 30.6\u25e6 and the pulley-fibre angle is \u03d5 = 146\u25e6). Drive exciting signals can be of a rectangular, a trapezoidal and a quasi-sinusoidal shape and there is a possibility of variation of a signal rate [10]. The amplitudes of the drive displacements are up to 90 mm. Time histories of the weight position (in the direction of the inclined plane; measured by means of a dial gauge), of the drive position (in vertical direction) and of the force acting in the fibre (measured on a force gauge at drive) were recorded using sample rate of 2 kHz."
        ]
    },
    {
        "image_filename": "designv11_64_0003602_1.5119067-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003602_1.5119067-Figure3-1.png",
        "caption": "Figure 3 Schematic of continuous coaxial powder nozzle for inside diameters and ultra-high-speed LMD.",
        "texts": [
            " As beam source a 4kW Laserline LDF-4000-8 diode laser source with beam converter is linked to the inside processing head via a 200 \u00b5m fiber. Since the nozzle position is fixed, the beam focal position is adjusted through positioning of the focusing lenses within the tubus. For the experiments, lenses forming a 1:1 image are used and arranged in a way such that the beam is focused above the working piece. In the working distance of approx. 8 mm from nozzle tip to working piece, a defocused, round, Gaussian laser spot with a diameter of approx. 1 mm is achieved, see Figure 3. With the continuous coaxial powder feeding nozzle, the powder gas stream is formed to the shape of a hollow powder gas cone. The powder nozzle can be adapted for different powder sizes by altering the distance of inner and outer cone through a threaded connection. For the used powder and a powder mass flow of 25 g/min a powder focus diameter of approx. 1 mm can be achieved. In Figure 3 the produced powder gas stream is depicted for powder mass flows of 5 g/min (left), 15 g/min (middle) and 25 g/min (right). The developed inside processing head is integrated into a Hornet Laser Cladding high-speed LMD system [8]. The system is based on a conventional lathe and retrofitted for ultra-high-speed LMD of rotationally symmetric components. The machine is fully enclosed by laser safe, light tight panels. One sliding door at the front side of the machine provides access to the work piece"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003048_ebccsp.2016.7605267-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003048_ebccsp.2016.7605267-Figure4-1.png",
        "caption": "Fig. 4. Quadrotor: fixed frame Ef = [ ef1 , e f 2 , e f 3 ] and body-fixed frame Eb = [ eb1, e b 2, e b 3]",
        "texts": [
            " The VTOL vehicle considered in the simulation is the well known four-rotor mini-helicopter so called quadrotor modeled as (6)-(7). Several simulations were carried out with a focus on two main outcomes: consensus and formation of VTOL vehicles. The quadrotor is a small aerial vehicle that belongs to the VTOL (Vertical Taking Off and Landing) class of aircrafts. It is lifted and propelled, forward and laterally, by controlling the rotational speed of four blades mounted at the four ends of a simple cross and driven by four DC Brushless motors (BLDC). On such a platform (see Fig. 4), given that the front and rear motors rotate counter-clockwise while the other two rotate clockwise, gyroscopic effects and aerodynamic torques tend to cancel each other out in trimmed flight. The rotation of the four rotors generates a vertical force, called the thrust T , equal to the sum of the thrusts of each rotor (T = f1+f2+f3+f4). The pitch movement \u03b8 is obtained by increasing/decreasing the speed of the rear motor while decreasing/increasing the speed of the front motor. The roll movement \u03c6 is obtained similarly using the lateral motors",
            " The yaw movement \u03c8 is obtained by increasing/decreasing the speed of the front and rear motors while decreasing/increasing the speed of the lateral motors. In order to avoid any linear movement of the quadrotor, these maneuvers should be achieved while maintaining a value of the total thrust T that balances the aircraft weight. In order to model the system\u2019s dynamics, two frames are defined: a fixed frame in the space Ef = [ e f 1 , e f 2 , e f 3 ] and a body-fixed frame Eb = [ e b 1 , e b 2 , e b 3 ], attached to the quadrotor at its center of gravity, as shown in Fig. 4. The components of the control torque vector \u0393 generated by the rotors are given by: \u03931 = dbm(um3 \u2212 um4) \u03932 = dbm(um1 \u2212 um2) \u03933 = km(\u2212um1 + um2 \u2212 um3 + um4) T = bm \u2211 l\u0304=1 4 uml\u0304 (28) with d being the distance from one rotor to the center of mass of the quadrotor. The specification and parameters of the quadrotor, used for the simulation, are given in the Table I. For this simulation, the control law (9) is implemented with control gains: \u03c11,2 = 4.2, \u03c11,2 = 1.74 and \u03ba = 0.075. For this propose, four agents are employed, each one represents a VTOL - UAV in simulation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001811_0954405415608784-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001811_0954405415608784-Figure6-1.png",
        "caption": "Figure 6. Loads acting on a ball.",
        "texts": [
            " According to the Pythagorean theorem, Figure 5 shows that cos ao = X2 fo 0:5\u00f0 \u00deD+ do sin ao = X1 fo 0:5\u00f0 \u00deD+ do cosai = A2 X2 fi 0:5\u00f0 \u00deD+ di sinai = A1 X1 fi 0:5\u00f0 \u00deD+ di 8>>>>>>< >>>>>: , fi = ri D , fo = ro D \u00f08\u00de where ao and ai are the contact angles for ball bearings of the outer and inner rings, respectively, and X2 1 +X2 2 = fo 0:5\u00f0 \u00deD+ do\u00bd 2 \u00f09\u00de A1 X1\u00f0 \u00de2 + A2 X2\u00f0 \u00de2 = fi 0:5\u00f0 \u00deD+ di\u00bd 2 \u00f010\u00de Consider the plane passing through the bearing axis with the centre of a ball located at any azimuth. If outer raceway control is approximated at a given ball location, the gyroscopic moment of the ball is resisted entirely by frictional force at the ball\u2019s outer raceway contacts, as shown in Figure 6, where the inner and outer ball friction force Ti =0 and To =2MG=D. The normal ball loads in accordance with normal contact deformations are as follows (Harris26) Qo(i) =Ko(i)d 1:5 o(i) \u00f011\u00de where Ko(i) is the load deflection coefficient of the ball bearings and do(i) is the contact deformation. For the steel balls, the centrifugal force, Fc, is Fc =2:263 10 11D3n2mdm \u00f012\u00de where nm is the ball orbital speed and dm is the pitch diameter. at The University of Auckland Library on June 5, 2016pib.sagepub.comDownloaded from From Figure 6, considering the equilibrium of forces in the horizontal and vertical directions To cos ao +Qo sin ao Ti cos ai Qi sin ai =0 \u00f013\u00de Fc +To sin ao +Qi cos ai Qo cos ao Ti sin ai =0 \u00f014\u00de where ai and ao are the inner and outer contact angles, respectively. Gyroscopic moment at each ball location is defined as follows (Harris26) MG = J vR v vm v v2 sin b \u00f015\u00de where vm is the orbital speed of the ball; vR is the rotational speed of the bearing element; J is the mass moment of inertia; and vR=v, vm=v, and b are described as follows, respectively (Harris11,26 and Jones10) tan b= sin ao cos ao + D dm \u00f016\u00de v vR = cos ao + tan b sin ao=1+ D=dm\u00f0 \u00de cos ao\u00f0 \u00de 1= D=dm\u00f0 \u00de cos b + cos ai + tan b sin ai=1 D=dm\u00f0 \u00de cos ai\u00f0 \u00de 1= D=dm\u00f0 \u00de cos b vm v = 1 D=dm\u00f0 \u00de cosai\u00bd 1+ cos ai ao\u00f0 \u00de\u00bd \u00f018\u00de To determine the values for da, the only remaining requirement is to establish an equilibrium condition for the entire bearing Fa = XZ j=1 To cos ao +Qo sin ao\u00f0 \u00de \u00f019\u00de where Fa and Z represent the axial external load and the number of balls, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.36-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.36-1.png",
        "caption": "FIGURE 8.36",
        "texts": [
            " The high-mobility multipurpose wheeled vehicle (HMMWV) example discussed in Chapter 1 is presented here, in more detail, to illustrate dynamic simulation and design. More than 200 parts and assemblies were created in the CADmodel, as shown in Chapter 5, Figure 5.23(a). The suspension was modeled in detail (Chapter 5, Figure 5.23(b)) such that the main aspects of the vehicle\u2019s performance Resultant shock travel distances for the current design (150 lb preload), 200 lb and 250 lb preload. could be captured accurately in motion simulation. A more detailed view of the front right suspension quarter is shown in Figure 8.36(a). A dynamic simulation model of 18 bodies and 21 joints, shown in Figure 8.36(b), was created and simulated in DADS with a total of 17 sec and a time step of 0.001 sec (Chang and Joo 2006). A 100 ft 100 ft terrain was used for simulation (see Chapter 5, Figure 5.24(b)). Note that the terrain was fairly bumpy. The maximum height of the bumps on the terrain was 7.68 in. The vehicle vibrated significantly toward the later stage of the simulation because of bumpy road conditions. In this model, the vehicle was \u201cdriven\u201d by a constant angular velocity of 1.53 rev/sec applied at the four wheels, which produced a path that went through both bumpy areas",
            " zwi(t) is the z-coordinate of the ith wheel center. Shock travel of the improved design by setting spring preload to 250 lb and changing the rocker height to 2.7 in. hi(t) is the height of the road profile corresponding to the ith wheel at the given time t. zds(t) and \u20aczds\u00f0t\u00de are the driver seat position and acceleration, respectively, in the z-direction of the global coordinate system (vertical). zch(t) is the z-displacement of the chassis with respect to the global coordinate system (shown in Figure 8.36(b)). Impact of adding slims to the camber angle. Note that in Eq. 8.106, ja(b) and jb(b) essentially characterize the deformation of the tires. ju simply represents the upper bound of the respective constraints j(b). Note also that the tire radius is 18 in. and zwi(t) 18 is 0 if no deformation occurs in the tire. jg(b) specifies the wheel center position with respect to the chassis in the vertical direction. The function in the integrand F(p(t)) of Eq. 8.106 is defined as follows (U.S. Tank-Automotive Research and Development Command, 1979), F p t \u00bc p1 t 0:108p4 t \u00fe 0:25p6 t p7 t (8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002611_sta.2015.7505116-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002611_sta.2015.7505116-Figure1-1.png",
        "caption": "Fig. 1. Illustration of a robot\u2019s parameters",
        "texts": [
            " k is a positive multiplier which delimits the coverage area of the force field. Dmax is the maximum active distance of a robot\u2019s force field and Dmin is the distance while the repulsive force is maximum. \u03c10 is a positive fractional number that heavily influences how close the robot can be separated from obstacles. Tp represents the priority of a task undertaken by the robot. For the case of a single robot, Tp is set 1. So, Dmax and Dmin depend on the parameters k, C, P, Q and \u03c10. Then, in our work, we use PSO in the optimization of these parameters. Figure 1 shows the different parameters used to compute the repulsive force. The repulsive force generated by a robot is defined by : |Frep\u2212rob| = \u23a7\u23a8 \u23a9 0 if D > Dmax P \u2217 Dmax\u2212D Dmax\u2212Dmin if Dmin \u2264 D < Dmax Fmax if D < Dmin (5) Where D is the shortest distance between point (x,y) and the perimeter of the robot. P limits the magnitude of the repulsive force and it\u2019s a positive constant scalar. The magnitude of the repulsive force changes from P to 0, when D changes from Dmin to Dmax. While D is less than Dmin, the magnitude of the repulsive force is equal to Fmax, the maximum repulsive force"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.2-1.png",
        "caption": "FIGURE 8.2",
        "texts": [
            " This analysis is necessary when torque or force is involved in motion analysisdfor example, the torque required to generate sufficient digging force for a backhoe. In those cases, the dynamic behavior of a mechanism is governed by Newton\u2019s laws of motion. Even though design questions vary and they often must be determined for specific applications of certain fidelity levels, there are a few basic questions that are common and typical for all motion applications. A single-piston engine, shown in Figure 8.2, is used to illustrate some typical questions: \u2022 Will the components of the mechanism collide or interfere in operation? For example, will the connecting rod collide with the inner surface of the piston or the inner surface of the engine case during operation? \u2022 Will the components in the mechanism move according to our intent? For example, will the piston stay entirely in the piston sleeve? Will the system lock up when the firing force aligns vertically with the connecting rod? \u2022 Howmuch torque or force does it take to drive the mechanism"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001920_jae-141922-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001920_jae-141922-Figure3-1.png",
        "caption": "Fig. 3. FE model for examining repelling force in the x-direction (Model I).",
        "texts": [
            " In this case, we selected the exciting current of EM3 at a 90-degree phase lag to the exciting currents of EM1 and EM2 so that the ac ampere force could be upwardly generated. Figure 2 illustrates a polarity pattern in our maglev system. Tooth pitch in each rail is 60 mm as a noted scale in Fig. 2. Here, we define the system when the electric circuit of EM3 is not used as ac induction type maglev method. In contrast, we also define the status when the exciting current of EM3 flows through the circuit as ac ampere type one. Figure 3 shows a FE model for examining repelling force in the conveyance (x) direction (Model I). Outer air domain and meshed elements are made invisible in the model. Levitation height of the Al plate Fig. 4. FE model for examining restoring force in the y-direction (Model II). from stator rails of EM1 and EM2 was set to 5 mm. The material property of each thin domain shown in Fig. 3 was assigned in either air or aluminum. The domain enables positional changes of the plate without deformation of meshed elements. Configurable maximum displacement of the plate is 144 mm in the x-direction. Similarly, Fig. 4 shows a FE model for examining restoring force in the y-direction (Model II). Configurable displacement of the plate is from \u22124 mm to 4 mm. Main parameters for FE analysis are listed in Table 1. I1 and I3 stand for the amplitude of exciting current, and j is the imaginary unit",
            " We can see that levitation force increases in almost direct proportion to the increase of current of EM3, and slightly increases when the Al plate moves in the y-direction. The result proves that ac ampere force contributes as a part of levitation force in the z-direction, and the Al plate receives restoring force with y- displacement. Figure 8 shows repelling forces when the Al plate moves in the x-direction. Each pair of exciting currents was selected as the value which were experimentally capable of levitating the plate with just a 5 mm height. Besides, the repelling forces, which were calculated from the Model I shown in Fig. 3, indicate the force change for one period of polarity pattern in Fig. 2. If repelling force is zero and slope of the curve is positive, it means an unstable point. The plate will slide to a stable point where a curve section with negative slope and the zero force line meet. In other words, repelling forces should act as resistive forces when an Al plate is conveyed from one stable point to the next one. The repelling forces in the ac ampere type method with the exciting current I3 of 2.75 A was almost half compared to that in the ac induction type one at inner peak points in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003105_eais.2015.7368805-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003105_eais.2015.7368805-Figure1-1.png",
        "caption": "Fig. 1. Conceptual presentation of Reynolds\u2019 Boids model. (a) Separation rule. (b) Alignment rule. (c) Cohesion rule.",
        "texts": [
            " A, G, E, and SR denote age, gender, energy level, and sight radius of the bird, respectively. C represents the center of mass of the closest flock to the bird in its sight radius, whose members belong to the same species as that bird belongs to. In order to simulate the choreographed motion of herbivore birds, we utilize five rules including three flocking rules proposed by Reynolds [6] (i.e. cohesion, alignment and separation) and two additional rules including escaping from a predator and moving towards food. Conceptual presentation of Reynolds\u2019 rules is shown in Fig. 1. Cohesion refers to steering towards the perceived center of mass of local flock-mates. The perceived center of mass is computed by (1). h i h j h i ij Sj h j p i SRtPtP| i ,NtPtC i \u2264\u2212\u2200=+ \u2211 \u2260 \u2208 |)()(:/))(()1( (1) Cp i(t) is the perceived center of mass for bird i, Ph i(t) is the position of herbivore bird i, SRh i is the sight radius of herbivore bird i, Si is the species that bird i belongs to, and N is the number of flock-mates of bird i that are in its sight radius and belong to the same species as bird i belongs to"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.10-1.png",
        "caption": "Fig. 3.10 Deformation of a beam in the x\u2013z plane with My(x) = const.",
        "texts": [
            "21), the distribution of stress over the cross section results in: \u03c3x(x, z) = +My(x) Iy z(x). (3.26) 3.2 Derivation of the Governing Differential Equation 99 100 3 Euler\u2013Bernoulli Beams and Frames The plus sign in Eq. (3.26) causes that a positive bending moment (see Fig. 3.4) leads to a tensile stress in the upper beam half (meaning for z > 0). The corresponding equations for a deformation in the x\u2013y plane can be found in [37]. In the case of plane bending with My(x) = const., the bending line can be approximated in each case locally through a circle of curvature, see Fig. 3.10. Therefore, the result for pure bending according to Eq. (3.25) can be transferred to the case of plane bending as: \u2212 EIy d2uz(x) dx2 = My(x). (3.27) Let us note at the end of this section that Hooke\u2019s law in the form of Eq. (3.20) is not so easy to apply12 in the case of beams since the stress and strain is linearly changing over the height of the cross section, see Eq. (3.26) and Fig. 3.9. Thus, it might be easier to apply a so-called stress resultant or generalized stress, i.e. a simplified representation of the normal stress state13 based on the acting bending moment: My(x) = \u222b\u222b z\u03c3x(x, z) dA, (3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure7.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure7.3-1.png",
        "caption": "Fig. 7.3 Configuration by X-ray imaging",
        "texts": [
            " In an X-ray generation tube, electrons are released by the source cathode and are accelerated toward the target anode in a vacuum under the potential difference ranging from 20 to 150 kV. When electrons collide with a metal target (anode), electromagnetic X-ray radiation can be generated due to the large drop in acceleration caused by the interaction between the anode and the electrons. During X-ray imaging, patients are asked to lie or stand between the X-ray emitter and a film which captures the X-ray image (Fig. 7.3). An X-ray image can effectively visualize the bone shape features of the patient, including any cracks and abnormal growth of bones. When the radiation beams enter the human body, they are either absorbed by tissues or penetrated through the body with different organs/tissues having various capabilities of absorbing X-rays. This is because they are made of different elements. Heavier atom, e.g., calcium in the bone, can absorb more X-rays as they contain more electrons. The radiation beam is attenuated in the human body due to the mass attenuation coefficients of physiological structures"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure14-1.png",
        "caption": "Fig. 14. Geometry of a Pivoted Leading Shoe",
        "texts": [
            " k = the distance in terms of drum radius from the centre of the brake drum to the pivot, or, in the case of a floating shoe, the perpendicular distance from the centre to the line of action of the abutment force. The assumptions made in the case of the pivoted shoe are that the line through the centre of pressure and the drum centre is at 90 deg. to the line through the pivot and the brake centre, and that the perpendicular distance from the pivot to the line of action of the shoe-tip load is equal to 2k x r, as shown in Fig. 14. . In the case of the sliding shoe, it is assumed that the shoe tip and the abutment forces are equidistant from the centre, as shown in Fig. 5, where OA = OB = h x r . The value of 1 depends on the included angle of the lining and is likely to lie between 1-1 and 1-17 (Barford 1933). The maximum value of k is limited by the necessity to house the pivot or abutment within the brake drum. It is diflicult for it to exceed 0.8, and 0-75 is a typical value. Leading shoes will \u201csprag\u201d or become self-applying when cot 0 or cosec 0 equals Elk, as the case may be",
            " 21 k+I sin0 heref fore F = t-(cosec e- l /k) ( l+- l 21 e - (cosec O--l/k)2\u2019 (e) Stepped cylinder, ratio s / l :- Let leading-shoe-tip load be x. Then trailing-shoe- tip load is ( s x x ) and X + ( S X X ) = 2. Therefore and 2 (S+l) 2s ( s x x ) = -- ( s+ l ) \u2019 2 21 2s x = - \u201c i = *{m) X(cosecB-Z/K)+(sx(cosec8+I/k) - -i (s + 1) 1. 21 (S + 1 ) msec 0 - (S - 1)2/k cosec2 8 -12/k2 (3) Factors for pivoted shoes with centre pressure at 90 deg. from the line joining pivot and drum centre, and the perpendicular distance from the pivot to the line of action of the shoetip load twice the distance from the pivot to the centre. In Fig. 14, which shows a leading shoe, from Acres, AB O N r A N F=-x--- ON -op = tan0 = p ; O r = ( l x r ) ; A B = 20A;= 2kr. 2Kr IrtanO F = - X r (kr-Zrtane) Therefore 2c cot e-l/k\u2019 - Similarly, for a trailing shoe, F = 21 cot B+l/k\u2019 A P P E N D I X I 1 ACRES, F. A. S. 1946 Proc. Inst. Automobile Eng., vol. 41, p. 19, \u201cSome Problems in the Design of Braking Systems\u201d. BARFORD, V. G. 1933 Proc. Inst. Automobile Eng., vol. 27, 52 INTERNAL EXPANDING SHOE CHASE, T. P. DAWTREY, L. H. 1930 Proc. Inst. Automobile Eng"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002357_978-3-319-17067-1_24-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002357_978-3-319-17067-1_24-Figure7-1.png",
        "caption": "Fig. 7 The experimental system",
        "texts": [
            " 5a, and the amplitude is 6 mm which meets the application requirements. The torque on the crank is shown in Fig. 5b. Figure 6a displays a cycle of steering simulation. And, the motor speed of the two quadruped mechanisms is set as 15 and 45r/min, respectively. Besides, Fig. 6b illustrates the lateral force on the leg mechanism which results in loosing connection and increasing friction between linkages. Therefore, the detailed mechanical design strategy of a revolute joint shown in Fig. 6c is adopted in each leg. As shown in Fig. 7, the experimental system composed of walking system, a control unit, a power supply, an operation unit is built (see Table 2 for a summary of the specifications of the DQV). See Fig. 8, the turning experiments are carried out to validate the feasibility of the experimental system and the locomotion of constructed prototype. The experiment illustrated in Fig. 9 testifies the zero radius steering ability of the DQV. Figure 10 shows the vertical obstacle passing experiments. The obstacle with the height of 20 mm is easy to overcome for the DQV"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000868_s12283-014-0166-y-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000868_s12283-014-0166-y-Figure2-1.png",
        "caption": "Fig. 2 Inertial coordinate system. The origin is at the center of the turning circle, with the XE-axis in the horizontal forward direction, the YE-axis in the horizontal lateral direction and the ZE-axis vertically downward",
        "texts": [
            " The pitching moment coefficient, CM, is positive (nose-up rotation) over almost the whole range. Beyond the stalling angle, |CM | is low between 32 and 40 . The dependence of CM on a was approximately linear in the range 0 \\ a\\ 25 . The standard deviation for CD is less than 0.03, except around the stalling angle, while those for CL and CM are less than 0.05 and 0.03, respectively. The irregularity (standard deviation) becomes large near the stall angle. The inertial coordinate system is shown in Fig. 2. The origin is at the center of the throwing circle, with the XEaxis in the horizontal forward direction, the YE-axis in the horizontal lateral direction and the ZE-axis vertically downward. The coordinate system in the discus body-fixed system is denoted by xb, yb, and zb (Fig. 3c). The origin is defined as the center of gravity of the discus. It is assumed that the geometric center of the discus coincides with the center of gravity, that its zb-axis is aligned with the transverse axis (axis of symmetry), and that xb and yb are aligned with the longitudinal axes in the discus planform",
            " The negative sign of ZE means the vertically upward direction, and the value of -1.6 is almost the highest release position for women [12]. The release height is generally 90 % of the thrower\u2019s height. 4.3 Constraint A constraint, g1, g1 \u00bc YLine XE tf YE tf [ 0 \u00f022\u00de is considered. This constraint means that the discus should make ground contact within the sector. Here, YLine(XE(tf)) is the sector boundary value of YE corresponding to XE (tf), which is defined by YLine XE tf \u00bc tan 34:92 2 XE tf \u00f023\u00de The angle of 34.92 is shown in Fig. 2. 4.4 Monte Carlo method In order to estimate F2 in Eq. (20), the flight distance should be simulated around FDc. There are two questions to consider in this process. The first question is how many flight simulations n in Eq. (20) are considered. The other is how wide the range for each control and design variable should be. For the first question, the higher the value of n in Eq. (20), the more convergent (constant) F2 will be, but the simulations will take a longer time to complete. It is possible to simulate FDi with variables chosen from a constant interval"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure5.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure5.2-1.png",
        "caption": "Fig. 5.2 Two-dimensional problem: plane stress",
        "texts": [
            " Using engineering definitions of strain, the following relations can be obtained [10, 17]: \u03b5x = \u2202ux \u2202x ; \u03b5y = \u2202uy \u2202y ; \u03b3xy = 2\u03b5xy = \u2202ux \u2202y + \u2202uy \u2202x . (5.1) In matrix form, these three relationships can be written as \u23a1 \u23a2 \u23a3 \u03b5x \u03b5y 2\u03b5xy \u23a4 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a3 \u2202 \u2202x 0 0 \u2202 \u2202y \u2202 \u2202y \u2202 \u2202x \u23a4 \u23a5 \u23a5 \u23a6 [ ux uy ] , (5.2) or symbolically as \u03b5 = Lu, (5.3) where L is the differential operator matrix. 5.2 Derivation of the Governing Differential Equation 243 5.2.2 Constitutive Equation 5.2.2.1 Plane Stress Case The two-dimensional plane stress case (\u03c3z = \u03c3yz = \u03c3xz = 0) shown in Fig. 5.2 is commonly used for the analysis of thin, flat plates loaded in the plane of the plate (x\u2013y plane). It should be noted here that the normal thickness stress is zero (\u03c3z = 0) whereas the thickness normal strain is present (\u03b5z = 0). The plane stress Hooke\u2019s law for a linear-elastic isotropic material based on the Young\u2019s modulus E and Poisson\u2019s ratio \u03bd can be written for a constant temperature as \u23a1 \u23a3 \u03c3x \u03c3y \u03c3xy \u23a4 \u23a6 = E 1 \u2212 \u03bd2 \u23a1 \u23a3 1 \u03bd 0 \u03bd 1 0 0 0 1\u2212\u03bd 2 \u23a4 \u23a6 \u23a1 \u23a3 \u03b5x \u03b5y 2 \u03b5xy \u23a4 \u23a6 , (5.4) or in matrix form as \u03c3 = C\u03b5, (5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure13.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure13.17-1.png",
        "caption": "Fig. 13.17 Characteristic of the variable voltage source in the projective coordinates",
        "texts": [
            "34) V0 \u00bc V1chc\u00fe I1 q shc \u00bc V1chc\u00fe P1 qV1 shc: This expression is similar to (13.6) and Fig. 13.7. We suppose the derivation equals zero; that is, dV0 dV1 \u00bc chc P1 q\u00f0V1\u00de2 shc \u00bc 0: From here, we obtain V1M \u00bc ffiffiffiffiffiffiffiffiffiffiffiffi P1 q thc s : \u00f013:39\u00de Therefore, we get the allowable minimum voltage source values V0M \u00bc 2 \u00f0chc\u00de ffiffiffiffiffiffiffiffiffiffiffiffi P1 q thc s \u00bc 2V1M chc: \u00f013:40\u00de We must show the single-valued area of the voltage source characteristic. To do this, we consider this characteristic in the projective coordinates in Fig. 13.17. The parallel tangent lines V0M intersect at the point S as the infinitely remote point 1 onto the axis I0. This point S is the pole and the straight line M \u00feM is the polar. Therefore, we get mapping of the \u201clower\u201d part of our curve onto the \u201cupper\u201d part. The points M \u00fe ;M are the fixed or base points. So, the obtained single-valued area involves the characteristic points V0M ;1; V0M and point V \u00fe 0 too. 378 13 Power-Source and Power-Load Elements In this case, the value H \u00fe is a scale value. The conformity of D1 is presented in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003666_s00022-015-0274-2-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003666_s00022-015-0274-2-Figure4-1.png",
        "caption": "Figure 4 A continuous motion by one moving edge, a an equilateral triangle T = abc with the center of gravity g and the midpoint h of ab, b the folded triangle Tp on pg obtained by creases for a mountain fold on pg and valley folds on {ga, gb, gc} and c the flat folded triangle Th on hg",
        "texts": [
            " A truncated regular tetrahedron Q has four regular hexagonal faces and four equilateral triangular faces. By Lemma 2.3, it is easy to see that the six hexagonal faces can be flattened continuously since they form a subset of the regular tetrahedron. We show how to flatten the four equilateral triangular faces in accordance to the continuous folding process for the rest of Q. We define two types of continuous folding processes for an equilateral triangle. Definition 3.1. Let T = abc be an equilateral triangle, g be the centroid of T , and h be the midpoint of ab (Fig. 4a). Apply valley folds on {ga, gb, gc} and mountain folds on gh until gh touches bgc. Then the resulting figure is flat and we call it a flat folded state of T on gh and denote it by Th. Let p be any point on bh. Fold T with a mountain fold on pg and valley folds on {ag, bg, cg} and attach pgb onto the face bgc. We call the resulting figure a folded triangle on pg and denote it by Tp (Fig. 4b). By moving p from b to h, the crease (edge) bg moves from bg to hg, and we obtain a continuous folding process from T to Th. We call such process a continuous motion by one moving edge. There is another continuous folding process from T to Th as follows. Consider T as the face abc of a regular tetrahedron and apply the continuous folding process shown in Lemma 2.3. Then T is folded continuously to Th by moving creases (edges) {aq, bq} where q moves from h to g (Fig. 5). We call such process a continuous motion by two moving edges",
            " It suffices to give a continuous folding process for those four triangular faces of Q in accordance to the continuous motion of Q\u2032. We denote by Ta, Tb, Tc and Td, those four triangular faces obtained by truncating P at vertices a, b, c and d respectively. Denote by g1 and g2 the centroids of abc and abd, respectively. We apply a continuous motion by one moving edge to Tb and Tc as follows. Let Tb = b1b2b3 as shown in Fig. 7. For each 0 \u2264 t \u2264 1, let pt be the intersection of b1b2 and rtb, and take pt as p in Fig. 4a. Then Tb is flat folded using a continuous motion by one moving edge consistent with the one for Q\u2032. Similarly, the triangle Tc is also flat folded using a continuous motion by one moving edge consistent with the one for Q\u2032. Let gb and hb be the centroid of Tb and the midpoint of b1a2. Since the line segments corresponding to h2g1 and h2gb have the same direction in the flat folded states of Q and Ta, we can flatten Tb in the way that is consistent to the motion of Q. We can also flatten Tc similarly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure4.20-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure4.20-1.png",
        "caption": "Fig. 4.20 a Efficiency of a two-port for load changes into the Cartesian coordinates, b this efficiency is as closed curve in projective coordinates",
        "texts": [
            "5) n11 \u00bc VOC 1 V1 1 VOC 1 VSC 1 \u00bc VS: From this, V1 1 \u00bc VOC 1 VS\u00f0VOC 1 VSC 1 \u00de \u00bc F\u00f0VS\u00de: V OC 1 V SC 1 I SC 0 , I1 0 , IOC 0 V 1 1 V S F V S Gen F -1 I 1 V 0 , I 0 Y 1 L1 (I 1 0 ) Y 1 L1 Y OC L1 Y SC L1 I SC 0 , I1 0 , I i 0 , IOC 0 Y i L1 Y 1 L1 Gen I 1 V 0 , I 0 we find the value Y1 L1 \u00bc YSC L1 \u00feNi L1Y OC L1 m1 L1 1\u00feNi L1m 1 L1 \u00bc Y1 L1\u00f0m1 L1\u00de; Ni L1 \u00bc YSC L1 Yi L1 Yi L1 YOC L1 : On the other hand, the cross ratio, by the input currents ISC0 ; I10 ; I i 0; I OC 0 , has the view m1 L1 \u00bc \u00f0ISC0 I10 Ii0 I OC 0 \u00de \u00bc I10 ISC0 I10 IOC0 Ii0 ISC0 Ii0 IOC0 \u00bc m1 L1\u00f0I0\u00de: Using this expression, we obtain the value Y1 L1\u00f0I0\u00de. 4.4 Deviation from the Maximum Efficiency of a Two-Port We use the results of Sects. 1.5.1 and 2.3. Let us consider a more complex case of a quadratic curve, as the efficiency of a two-port in accordance with expression (1.34) KP \u00bc KG 1 KG A KG : \u00f04:49\u00de This expression represents a hyperbola in Fig. 4.20 for all area of load changes. The positive load consumes energy; the maximum power transfer ratio (1.33) K \u00fe P \u00bc \u00f0 ffiffiffi A p ffiffiffiffiffiffiffiffiffiffiffi A 1 p \u00de2: \u00f04:50\u00de Then, the corresponding voltage transfer ratio K \u00fe G \u00bc A ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi A\u00f0A 1\u00de p : \u00f04:51\u00de In turn, the negative load returns energy and we get the corresponding maximum values K P \u00bc \u00f0 ffiffiffi A p \u00fe ffiffiffiffiffiffiffiffiffiffiffi A 1 p \u00de2; K G \u00bc A\u00fe ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi A\u00f0A 1\u00de p : \u00f04:52\u00de Next, we must determine all the characteristic points [6, 11]. Obviously, there are points B1; 0;K P ; 1;A1;K \u00fe P by Fig. 4.20b. These points correspond to points T \u00bc 1; 0;K G ; 1;A;K \u00fe G of the axis KG. 120 4 Two-Port Circuits It is possible to take up the points 0; 1 as the base points and the point K G as a unit point. But visibly, the other characteristic points have to be defined relatively to these basic points and not depend on the parameter A of comparable two-ports. Therefore, we must, at first, define the possible systems of all the characteristic points. To do this, we will study a regime symmetry of efficiency using the regime symmetry for the load power of Sect. 2.3. 4.4.1 Regime Symmetry for the Input Terminals In Fig. 4.20b, the pole S and polar TA1 determine the mapping or symmetry of the region of power delivery by the voltage source V0 (above of the polar) on the region of power consumption of this voltage source (below of the polar). The point K \u00fe P passes into the point K P . Points B1;A1 are the fixed base points and KG\u00f0B1\u00de \u00bc 1, KG\u00f0A1\u00de \u00bc A. A hyperbola point is assigned by the hyperbolic rotation of radius-vector RSS. 4.4 Deviation from theMaximum Efficiency \u2026 121 In turn, the pole T and polar SM determine the mapping or symmetry of the hyperbola relatively to the straight line K \u00fe P K P or to the points of maximum efficiency",
            " The maximum output voltage at OC output, when I1 = 0, VOC 1 \u00bc Y10 Y11 V0: The maximum output power of the Th\u00e9venin equivalent generator PGM \u00bc VOC 1 ISC1 \u00bc \u00f0Y10\u00de2 Y11 \u00f0V0\u00de2: \u00f04:64\u00de In turn, the known value, as the effectiveness parameter, A \u00bc ch2c \u00bc P0M PGM : \u00f04:65\u00de 126 4 Two-Port Circuits Taking into account expressions (4.62)\u2013(4.65), we obtain the following equation P1 P0M \u00bc 1 P0 P0M A 1 P0 P0M 2 : It is possible to consider P0M \u00bc 1. Therefore, P1 \u00bc 1 P0\u00f0 \u00de A 1 P0\u00f0 \u00de2: \u00f04:66\u00de As it was told, expression (4.49) represents the hyperbola in Fig. 4.20 with the center S. This center coordinate corresponds to the value KPS \u00bc 2A 1: \u00f04:67\u00de Using (4.50) and (4.52), we get the following equalities K \u00fe P \u00bc KPS 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi A\u00f0A 1\u00de p ; K P \u00bc KPS \u00fe 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi A\u00f0A 1\u00de p : \u00f04:68\u00de Therefore, K \u00fe P \u00feK P \u00bc 2KPS: 4.5.2 Parallel Connection of Two Converters Let us consider two regulable converters or voltage regulators VR1;VR2 with a common load R1 in Fig. 4.25a. Two-ports TP1; TP2 with the corresponding effectiveness parameters A1; A2 are losses of these regulators are shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002698_978-981-10-1602-8_7-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002698_978-981-10-1602-8_7-Figure2-1.png",
        "caption": "Fig. 2 A schematic drawing of the experimental sample dimensions for the three-point flexure romax calculation and setting of the fatigue test",
        "texts": [
            " The material used in this study is a Ni-based superalloy INCONEL 718 with a chemical composition (in wt%) shown in Table 1. The material was heat treated, according to the supplier\u2019s BIBUS Ltd. (CZ) material sheet, at 980 \u00b0C/1 h AC (air cooled) and heating at 720 \u00b0C/8 h followed FC (furnace cooled) (50 \u00b0C per hour) to temperature 620 \u00b0C holding time 8 h and air cooled. The achieved mechanical properties of the material with grain size ASTM 12 are in Table 2. The experimental material for the three point flexure fatigue test was machined down into a simple blocky shape as reported in the schematic drawing (see Fig. 2). Also, for the calculation of the maximum bending stress, we used the formula (1): romax \u00bc 3 F L 2 b h2 MPa \u00f01\u00de where romax is the maximum bending stress (MPa), F is the dynamic load (N), L is the distance of supports (mm), b is the specimens width (mm), and h is the sample height (mm). The three point flexure fatigue test was carried out on testing machine ZWICK/ROELL Amsler 150 HFP 5100 at room temperature with a static pre-load Fstatic = \u221215 kN (this value may be considered as Fmedium when dynamic load changes from maximum to minimum around this Fmedium) and dynamic force Fdynamic varies from 6",
            " The S-N curve was plotted from the measured values, which gives the relation between the maximum bending stress romax and the number of cycles Nf. Fractography analysis of broken samples was also carried out. For fractography analysis we used the scanning electron microscope TESCAN Vega II LMU. All fractography analyses were carried out in order to describe mechanisms of fatigue crack initiation, fatigue crack propagation and final static fracture of the samples. For low-frequency fatigue testing we used 10 samples with simple blocky shape as seen in Fig. 2. Figure 3 shows the S-N curve of IN718 obtained from the three-point flexure fatigue tests at room temperature with a frequency of 150 Hz under the load ratio of R = 0.11. Obtained results were approximated using Eq. (2), which is Basquin\u2019s formula for the S-N presentation and approximation. This approach was also used for other materials [24]. ra \u00bc 2591 N 0:0811 f \u00f02\u00de where rf\u2032 = 2591 is the coefficient of fatigue strength, and \u22120.0811 = b is the lifetime curve exponent. It is clearly seen from the measured S-N curve that the fatigue life increases with decreasing stress amplitude and the S-N curve appears to continuously decline as the life extends"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002383_s13534-015-0188-9-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002383_s13534-015-0188-9-Figure5-1.png",
        "caption": "Fig. 5. Stress distribution obtained from finite element analysis for evaluation of structural stability of RehabWheel.",
        "texts": [],
        "surrounding_texts": [
            "Figs. 5 and 6 indicated the verification of the structural stability of the designed system and the control operation verification of the gait assistive function. The results of the structural stability of the system confirmed the occurence of von Mises stress below 88.6 MPa. This stress distribution was lower than the yield stress (276 MPa) of the aluminum material (AL6061-T6) used to construct the system; moreover, it was lower than the safety stress (92 MPa) applied with a safety factor of 3.0 for a conservative perspective. These results signify that the structural stability of the system during RehabWheel gait wil be suficient. The verification results of the control operation of the system confirmed that the exoskeleton motion patern of the RehabWheel, controled by artificial pneumatic muscles, accurately simulated joint angle change paterns that occur during general normal gait, within an eror range of 2.6 \u00b1 10.8%. These results signify that the control action of the RehabWheel using the pneumatic artificial muscles can be eficiently operated."
        ]
    },
    {
        "image_filename": "designv11_64_0000968_20140824-6-za-1003.01643-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000968_20140824-6-za-1003.01643-Figure5-1.png",
        "caption": "Fig. 5. The Decision Point This is the ultimate point where the UAV can make a loitering circle without the risk of colliding with the UAVs already on the objective circle. After the Decision Point, if the approaching UAV starts a loitering maneuver, it can potentially collide with the UAVs on the objective circle. As a result, at the Decision Point, the UAV must make the decision whether it can approach the objective circle without the risk of collision. It is clear that the Final Cruise Point is closer than the Decision point to the objective circle.",
        "texts": [
            " persistent loitering maneuvers can occur while the UAVs are in their Cruise state. 3.2.2 Collision avoidance in the Final Approach state In the Final Approach state, possible collisions between the approaching UAVs and the ones already moving on the objective circle must be avoided. The collision avoidance scheme presented in the following will prevent deadlocks, i.e. persistent loitering maneuvers in the Final Approach state. Here, a definition is made: Definition 6: The Decision Point, shown in Fig. 5, is a point on the cruise segment of the flight where the minimum distance between the center of the right loitering circle and that of the objective circle is UAVlo rRR 2++ . For an approaching UAV, the length of the final approach curve is computed for a given set of { }UAVlo rRR ,, . Here, a conservative solution is considered. The approaching UAV starts the Final Approach state only if there is no UAV on the objective circle in the interval [ ]arr\u03b1\u03b1 min . As defined before, arr\u03b1 corresponds to the angular position of the Arrival Point at the objective circle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002748_ijmic.2016.075271-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002748_ijmic.2016.075271-Figure11-1.png",
        "caption": "Figure 11 Actual position tracking path within desired region (see online version for colours)",
        "texts": [
            " Figure 6 shows the tracking errors in x, y and z direction is converging to zero. Figure 7 shows orientation angle roll (\u03c6), pitch (\u03b8) and yaw (\u03c8) to be regulated to constant desired orientation. The control torque input is depicted in Figure 8 for position tracking while Figure 9 shows the control input evolution for orientation regulation. The estimated parameters 1\u0398\u0302 and 2\u0398\u0302 are depicted in Figure 10. This figure shows that the estimated parameter values remain bounded for the simulation period. Finally, Figure 11 shows the actual tracking path of the vehicle position in 3D operational space. The blue line shows the vehicle trajectory within the eight-shaped desired region. The initial position was outside the desired region, but as time evolves, the actual path continues to remain within the region. Evolution of PE with time is shown in Figure 12 for region tracking and tracking purpose respectively. Furthermore, for comparison purpose, time evolution of kinetic energy (KE) for conventional tracking and region tracking problem has also been shown in Figure 13"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003763_insi.2015.57.5.283-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003763_insi.2015.57.5.283-Figure1-1.png",
        "caption": "Figure 1. Spur gear system model",
        "texts": [
            " A dialogue between analytical equations solved numerically and empirical experiments is the basis for this work. The scope of this work is primarily a validation of the dynamic model. As a secondary objective, the fault detection capabilities of features in the order domain will be evaluated. A dynamic model of a spur gear transmission, which includes eight degrees of freedom, is developed in order to describe the dynamic vibration response of the experimental gearbox system. The model incorporates all the components of the existing system, as presented in Figure 1. \u03b8i (i = g, p, m, b) is, respectively, the torsional displacement of the gear, the pinion, the motor and the brake wheel, all of which are assumed to be rigid bodies. xj and yj (j = 1, 2) are, respectively, the transverse displacements of the gear and the pinion. The shafts are assumed to be of finite torsional rigidity and supported by bearings, which are modelled by linear springs and viscous dampers[3]. Tm(t) is a torque input and Tb(t) is a torque output, all of which are generally composed of constant and harmonic components"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002215_6.2014-3261-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002215_6.2014-3261-Figure1-1.png",
        "caption": "Figure 1. The cyclogyro developed by us",
        "texts": [
            " The aero-elastic analysis also showed that the torsion of the blade will cause lower thrust of the cycloidal propeller. They also developed a cyclogryo which can fly in any direction7-11. Kan Yang made CFD analysis incorporated with the aero-elastic model developed by Moble Benedict. The CFD analysis signified that both 2D and 3D analysis could produce results comparable to experiment results12. The experimental studies and the numerical simulations are also carried out by us and several successful cyclogyroes had been developed by us13,14,15, as shown in Fig. 1. C D ow nl oa de d by K U N G L IG A T E K N IS K A H O G SK O L E N K T H o n Ja nu ar y 14 , 2 01 6 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /6 .2 01 4- 32 61 From the numerical simulations, it is found that the motion of the cycloidal propeller blade is comprised of the rotation about rotor shaft and the pitching oscillation about the pitching axis on the blade. If viewed in the moving reference frame attached to the cycloidal propeller, each blade is actually performing pitching oscillation in the curved flow"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001062_s00707-013-1041-9-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001062_s00707-013-1041-9-Figure9-1.png",
        "caption": "Fig. 9 The phase planes {y\u2032(t); y(t)} corresponding to 21 periods of stiffness function C(t) of variant solution (x1) for p = 1 (a), p = 6 (b) at frequency tuning \u03bds = 2.0 and at quadratic time variable damping 2kv(t)",
        "texts": [],
        "surrounding_texts": [
            "The work forms a brief introductory preexperimental study of basic research of the influence of lightening holes in the discs of gears on the internal dynamics (i.e. time heteronymous damping of relative motions y(t) in the gear mesh) in the high-speed lightweight planetary gear systems, such as they occur in the aerospace industry with turboprop drive units.3 Complexity, i.e. multiplication, of resonance bifurcation characteristics and their stability requires necessary theoretical analytical numerical solutions of weakly and, at impact effects in the gear meshes, strongly 3 Turbine revolutions for such drive units are currently up to 90,000 rpm. nonlinear systems with many degrees of freedom. The problems of internal dynamics including the determination of formation conditions of irregular chaotic motions at internal high frequency excitation by functions expressed by Fourier\u2019s series will be therefore solved analytically by means of the theory of transformation of differential boundary value problems to equivalent nonlinear integro-differential equations with solving kernels of Green\u2019s type and by the method of successive approximations. This study constitutes the partial methodological analysis of the influence of possible quadratic time variable damping in the gear mesh on the internal dynamics of weight lightened cog wheels in transmission systems from the point of view of the quality of damping coefficients in the gear meshes. The influences of all possible variants and combinations of linear and nonlinear constant and time variable damping forces will be subjected to deeper qualitative and quantitative analysis in the next studies so that the whole damping characteristics obtained by experimental methods could be simulated and approximated by means of theoretically modified damping characteristics of the solved systems. This work is also an introductory study for deeper dynamic research of the theory of elastohydrodynamic lubrication in gear meshes at strong nonlinearities arising from the impact effects and the rolling and sliding motions of gear profile flanks. The study of the influence of \u201czero\u201d approximation of the time heteronymous damping in gear meshes of transmission systems has generally to contribute to the progressive direction of beginning experimental research in a given field. Acknowledgments This study has been elaborated with institutional support RVO//: 61388998 in the Institute of Thermomechanics AS CR, v.v.i. Prague."
        ]
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure9.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure9.3-1.png",
        "caption": "Fig. 9.3 Material wear and volumetric loss",
        "texts": [
            " Roughly speaking, when the shearing stress is beyond the shear strength of either material of the interface, the material bodies can slide on each 250 9 Biocompatible Material Selection other. Such stress level is sufficiently high to cause the interfacial materials to gradually become damaged as the shear stress is sufficiently high. Furthermore, considering that most engineering surfaces are rough, the real contacting surfaces between single asperities and the material particles come off from the damaged asperities. The loss of material volume VW from one surface due to grinding between two material surfaces along a grinding distance LW under a normal load F (Fig. 9.3) has been characterized by the wear effect with the steady-state wear equation: VW \u00bc KW FLW 3HB \u00f09:1\u00de where KW (which is a dimensionless quantity) is the standard wear coefficient of the surface with the volume loss VW; andHB is the Brinell hardness for the surface of the softer material surface. Notably, the volume loss can occur on both sides of the grinding interface. To a certain extent, the ratio F/HB reflects the real contact area between the materials as a softer material would deform further on the asperities of the hard material surface with a larger contact area"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000288_978-3-642-40849-6-Figure2.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000288_978-3-642-40849-6-Figure2.1-1.png",
        "caption": "Fig. 2.1. The lumped parameter model of the thrusting system",
        "texts": [
            " The force transmission behavior of the thrusting system with different grouping strategiesis analyzed. A load equivalent model for cutterhead which contains different kinds of cutters is given. Based on above analysis, a simulation model is built by using MATLAB/SIMULINK. The dynamical behavior of the thrusting system with different grouping strategies is investigated numerically. In this section, the dynamics model of redundant thrusting system in shield machine is proposed by the Newton-Euler method. Then grouping strategy is introduced to form a new dynamics formulation. Figure 2.1 shows the lumped parameter model of the thrusting system. Origin O of fixed coordinate system O xyz\u2212 is located at the geometric center of fixed platform and its x axis is vertical to the platform. The float coordinate system ' ' ' 'O x y z\u2212 and the fixed coordinate system O xyz\u2212 are connected by hydraulic cylinders. Axes of fixed coordinate system and float coordinate system are parallel at the initial moment, respectively. When the thrusting system works in the tunnel, the thrusting system only has three degrees of freedom: translation along tunneling direction x axis, rotations around the y axis and the z axis in the excavated section"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000474_0954407015601265-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000474_0954407015601265-Figure2-1.png",
        "caption": "Figure 2. Three possibilities for identifying for identifying the longitudinal transmission behaviour of the suspension: case a, bump sensitivity; case b, longitudinal transmission behaviour of the suspension with slip stiffness of the tyre; case c, longitudinal transmission behaviour of the suspension without slip stiffness.",
        "texts": [
            " The forces and accelerations measured on a real road are compared with those measured on the suspension test rig. Therefore, two bumps (of heights 10 mm and 30 mm) are investigated on a real road. The car is driven over these bumps with different velocities. For the 30 mm bump, velocities from 10 km/h to 40 km/h are used and, for the 10 mm bump, the velocities were in the range from 10 km/h to 80 km/h in 10 km/h steps. On the suspension test rig, only one bump is reconstructed. This bump corresponds to a 30 mm bump and a velocity of 20 km/h on a real road (Figure 2, case a). Case a. This consists of a push against the tyre in the vertical direction and the longitudinal direction to reconstruct the bump sensitivity. In the second part the longitudinal transmission behaviour of the suspension is investigated. For this purpose, the car is driven over a reference track with a constant velocity. This reference track represents a typical uneven road profile such as found for middle European roads. The results of this measurement are compared with the results on the dynamic suspension test rig. On the suspension test rig, two measurement set-ups are possible (Figure 2, case b and case c). Case b. The pulse is in the longitudinal direction against the wheel carrier and is transmitted by a fixed at UNIV CALIFORNIA SANTA BARBARA on February 25, 2016pid.sagepub.comDownloaded from connection to the cylinder at the height of the wheel centre to identify the longitudinal transmission behaviour of the suspension in combination with the slip stiffness of the tyre. Case c. The pulse is in the longitudinal direction against the wheel carrier in the wheel centre with a mounted dummy wheel which is also connected to the vertical cylinder to adjust the wheel load"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.12-1.png",
        "caption": "FIGURE 6.12",
        "texts": [
            " (3 degree-of-freedom) linear model to assess the influence of suspension characteristics on straight line stability. These models belong very firmly in the \u2018analysis\u2019 segment of the overall process diagram described in Chapter 1, Figure 1.6. The functional representation of the model is based on components that describe effects due to kinematics dependent on suspension geometry and also elastic effects due to compliance within the suspension system. A schematic to support an explanation of the function of this model is provided in Figure 6.12. If we consider first the kinematic effects due to suspension geometry we can see that there are two variables that provide input to the model: \u2022 Dz is the change in wheel centre vertical position (wheel travel) \u2022 Dv is the change in steering wheel angle The magnitude of the wheel travel, Dz, will depend on the deformation of the surface, the load acting vertically through the tyre resulting from weight transfer during a simulated manoeuvre and a representation of the suspension stiffness and damping acting through the wheel centre"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001690_detc2015-46173-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001690_detc2015-46173-Figure1-1.png",
        "caption": "Figure 1. Shell element parameterizations",
        "texts": [
            "org/pdfaccess.ashx?url=/data/conferences/asmep/86618/ on 02/27/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use laminated composite material model is implemented in shear deformable shell element based on the absolute nodal coordinate formulation with the transverse slope coordinates for modeling the fiber-reinforced rubber (FRR) structure of a tire model. Furthermore, a physics-based tire model is developed for use in multibody dynamics simulation of ground vehicles. As shown in Fig. 1, the global position vector ir of a material point T[ ]i i i ix y zx in shell element i is defined as ( , ) ( , ) i i i i i i i i m i x y z x y z     r r r (1) where ( , )i i i m x yr is the global position vector in the middle surface and ( , )i i i ix y z r is the transverse gradient vector used to describe the orientation and deformation of the infinitesimal volume in the element. The position vector in the middle surface and the transverse gradient vector are approximated as follows: ( , ) ( , ) , ( , ) ( , ) i i i i i i i i i i i i i i m m p m gi x y x y x y x y z     r r S e S e (2) where i mS is the shape function matrix",
            " In the Kirchhoff plate theory, the plane stress is assumed and the in-plane strains of a plate [ ]T p xx yy xy  \u03b5 are defined as 0 p p z \u03b5 \u03b5 \u03ba (11) where 0 0 0 0[ ]T p xx yy xy  \u03b5 is a vector of the in-plane strains in the middle plane, and [ ]T x y xy  \u03ba is a curvature vector associated with bending and twisting. Using a linear orthotropic constitutive law, the in-plane stress vector is related to its strain vector by p p p\u03c3 C \u03b5 , and the material moduli is given as [13] 1 T p p  C R C R (12) In the preceding equation, the transformation matrix R is a function of the fiber angle  that defines the orientation of the fiber coordinate system o-12 with respect to the material frame o-xy of the plate as shown in Fig. 1. This matrix is defined by 2 2 2 2 cos sin sin 2 sin cos sin 2 sin 2 sin 2 cos 2                   R (13) and pC is the material moduli of an orthotropic material in the fiber coordinate system as 1111 1122 1122 2222 1212 0 0 0 0 p C C C C C          C (14) where 1111 1 12 21(1 )C E    , 2222 2 12 21(1 )C E    , 1122 21 1 12 21(1 )C E    , and 1212 12C G . While the coupling terms between the normal and shear strains in the fiber coordinate system are zero as observed in Eq. 14, the extension and shear coupling occurs for the stress and strain field defined in the material frame and the coupling terms in the material moduli matrix pC of Eq",
            " Two Gaussian integration points are used along the thickness when the elastic forces of each layer are evaluated. Orthotropic Saint-Venant-Kirchhoff Material Model For an orthotropic Saint-Venant-Kirchhoff material, the material moduli 2 /ijkl ij klC W      of an orthotropic lamina in the material frame is defined as [14] ( )( )( )( )ijkl i j k l abcd a b c dC C    b a b a b a b a (20) where 1 1 2 3( ) [ ]i  J b b b and abcdC is the tangent material moduli defined using 9 material parameters in the fiber coordinate system 1 2 3[ ]a a a as shown in Fig. 1, where the direction of fiber is defined along the coordinate 1. The material moduli abcdC in the fiber coordinate system are given as follows [14]: 1111 1122 1133 1122 2222 2233 1212 1133 2233 3333 2323 1313 0 0 0 0 0 0 0 0 0 0 0 [ ] 0 0 0 0 0 0 0 0 0 0 0 0 0 ijkl C C C C C C C C C C C C C                     (21) Mooney-Rivlin Material Model For modeling incompressible materials such as rubbers, Mooney-Rivlin material model is widely used. The energy density function is defined as [15] 2 1 1 2 2( 3) ( 3) ( 1) 2 K W C I C I J      (22) where 1C and 2C are material constants, 1 3 1 1 3/ ( )I I I , 2 3 2 2 3/ ( )I I I and 1 2 3( )J I , where I1, I2 and I3 are invariants of right Cauchy-Green tensor [15]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003134_978-981-10-2875-5_70-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003134_978-981-10-2875-5_70-Figure5-1.png",
        "caption": "Fig. 5 Two parts of the tetrahedral element",
        "texts": [
            " Then the whole deployable truss antenna shown in Fig. 1 can be obtained by extending the minimum composite units from the central axis of the antenna. The reflecting surface of the antenna spliced by the minimum composite units is shown in Fig. 4, in which the triangle planes are the undersurface of the tetrahedral elements and the color hexagons represent the minimum composite units. In order to analyze the DOF of the tetrahedral element, it is firstly divided into two parts in this section, as shown in Fig. 5. Then, the DOF of the first part is derived Fig. 4 Reflecting surface of the antenna spliced by the minimum composite units based on the reciprocal screw theory. Finally, the DOF of the tetrahedral element can be obtained after considering the constraint influence of the second part. The first part of the tetrahedral element can be regarded as a parallel mechanism (PM) composed of a fixed node A, a moving node B, and two supporting limbs. One limb connects the node B to the node A by three R joints (i.e., S1, S2 and S3), denoted by the limb RRR. The other limb consists of a closed-loop kinematic chain containing seven R joints (i.e., S4, S5, S6, S12, S13, S11 and S10) and a serial chain RR (i.e., S14 and S15), denoted by the limb (7R)-RR. A fixed coordinate frame Axyz is attached to the center of the node A, with x-axis perpendicular to the BA, yaxis pointing along the BA, and z-axis determined by the right-hand rule, as shown in Fig. 5. The closed-loop kinematic chain in the limb (7R)-RR can also be treated as a PM containing a fixed node A, a moving node H, two supporting limbs, RR (i.e., S10 and S11) and RRRRR (i.e., S4, S5, S6, S12 and S13). The constraint wrenches imposed on the node H by the limbs RR and RRRRR can be expressed in the A-xyz as ( ) ( ) ( ) ( ) ( ) T r1 T r2 4 4 T r3 11 10 11 10 11 10 10 11 11 10 11 10 10 11 10 11 11 10 T r4 4 4 4 10 4 10 10 4 4 10 T r5 5 2 5 2 5 2 2 5 0 0 0 0 0 1 0 0 0 0 0 0 0 0 H H H H H b a x x y y z z y z y z x z x z x y x y a b b z a z x b a y b c a c a b a b \u23a7 / = \u23aa \u23aa / = \u2212 \u23aa\u23aa / = \u2212 \u2212 \u2212 \u2212 \u2212 \u2212\u23a8 \u23aa / = \u2212 \u2212\u23aa \u23aa / = \u2212 \u2212\u23aa\u23a9 S S S S S , \u00f01\u00de where, a2 b2 c2\u00f0 \u00de, a4 b4 0\u00f0 \u00de and a5 b5 0\u00f0 \u00de denote the direction vectors of the S4, S10 and S12, x10 y10 z10\u00f0 \u00de and x11 y11 z11\u00f0 \u00de represent the position vectors of the centers of the joints S10 and S11, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001881_s1052618815030243-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001881_s1052618815030243-Figure1-1.png",
        "caption": "Fig. 1.",
        "texts": [
            " The critical values of the dissipation parameters were received higher than that at which the sufficient conditions of dynamic stability are observed. As the system loses its most significant nonstationary properties, such conditions can be interpreted as the qua sistationarity conditions. But the critical level of necessary dissipation can be rather high, and sometimes even unrealizable, which is why the search for other methods of performance of the quasistationarity con ditions is interesting. Below one of such methods is used as to the cyclic systems schematized in the form of the dynamic lattice. 2. In Fig. 1a two mechanisms are shown connecting the main shaft with the actuating devices schema tized in the form of a dynamic model given in Fig. 1b. Let us assume the following reference designation: Ji\u2014inertia moment; ci\u2014rigidity coefficients; \u03c8i\u2014dissipation coefficients; \u03a0\u2014operator corresponding to the function of the cyclic mechanism position \u03d52j = \u03a0j(\u03d51j), where \u03d51j, \u03d52j\u2014rotation angles of the input and output link; j\u2014mechanism number. In the general case this model reflects on of the modules of the DOI: 10.3103/S1052618815030243 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 44 No. 4 2015 QUASISTATIONARITY OF DYNAMIC MODES IN CYCLIC MECHANISMS 313 dynamic model of the circular structure of the multisection drive (Fig. 1c); here s = 1, \u2026, n\u2014number of sections [1, 3, 4, 7]. The nonlinear function of the position near the programmed motion is transformed with adequate accuracy into the characteristic of nonstationary constraint of type \u03a0(\u03d5* + qi) \u2248 \u03a0(\u03d5*) + \u03a0'(\u03d5*)qi, where \u03a0'(\u03d5*) = d\u03a0/d\u03d5 is the first geometric transfer function, \u03d5* = \u03c9t. For the purpose of better visualization and to reveal the physical presuppositions for implementation of the quasistationarity conditions let us first define the frequency characteristics of a single module (smax = n = 1) on the assumption of solid drive (c0 = \u221e), rigid mechanism (c = \u221e) and absence of con straints with other subsystems (\u0394c1 = 0, \u0394c2 = 0)",
            " Form coefficients also remain permanent within the whole kinematic cycle: Violation of the condition \u03bc = \u03b6 is illustrated by the curves 2 (\u03bc = 1, \u03b6 = 2) and 3 (\u03bc = 4, \u03b6 = 1) on the graph. It is interesting whether the revealed property will be also kept for a model of a more complex structure. The methods of dynamic calculation and analysis of such models based on the theory of regular systems are studied in detail in the monograph [7]. In the frame of this article for illustration let us confine our selves only to some results obtained for the model shown in Fig. 1c. The following initial data are assumed: q A0 pt \u03b1+( ).sin= 2 p1 k1 0.5 3 5\u2013( ), p2 k1 0.5 3 5+( ).= = \u03b21 0.5 5 1\u2013( )= , \u03b22 0.5\u2013 5 1+( ).= JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 44 No. 4 2015 QUASISTATIONARITY OF DYNAMIC MODES IN CYCLIC MECHANISMS 315 J1 = 0.05 kg m2; c1 = c0 = 2 \u00d7 103 N m; r0 = 0.75; \u03bc = \u03b6 = 0.5; n = 6. Let us note that this model as compared to the above studied model represents a dynamic lattice consisting of six sections in which the elastic char acteristics of the mechanisms are taken into account",
            " While defin ing the steady mode it is possible to use the numerical analytical method specified in [7] whose use permits limiting by the period 2\u03c0 upon direct summing of the oscillations, and implementing the periodicity con ditions analytically. In the initial system of the generalized coordinates the oscillations are described as q = S\u03b7. Let us illustrate some dynamic effects connected with implementation of quasistationarity conditions on the example of two section dynamic model upon taking into account the elastic dissipative character istics of the main shaft and actuating device (Fig. 1c). The analysis of the formula (12) shows that the persistence of natural frequencies in the absence of dis sipation despite the similar situation upon the natural oscillations of the systems with the constant param eters does not lead to the constant values of oscillations amplitudes Br = const. It is connected with the presence of gyroscopic forces. However a certain positive effect is observed which is especially obvious while analyzing accelerations. In Fig. 3 the function graphs are given for two cases: in the absence of dissipation and when it is taken into account ( = \u03b40)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002699_1.4034274-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002699_1.4034274-Figure1-1.png",
        "caption": "Fig. 1 The model (b) of the tested element (a) assumed in this paper",
        "texts": [
            " The sum of these functions F\u00f0x; _x\u00de \u00bc Fs\u00f0x\u00de \u00fe Fd\u00f0 _x\u00de described the total reaction force of the tested material component acting on the mass m, and p(t) is an external excitation force of any form. In this study, it was assumed that the tested element is made of a material with atypical (complex) rheological properties. Determining these properties can be carried out by analyzing the dynamic behavior of the concentrated mass m attached to any complex vibrating dynamical system by means of the tested material element (Fig. 1(a)). The structure as well as parameters of this complex system need not be known. Also, the location and direction of the applied excitation forces that affect this system do not need to be known. The analysis was performed based on the expanded rheological model of the Zener type. This extension is to replace the linear spring with the nonlinear spring described by the function S(x) of any form (Fig. 1(b)). As shown by adopting such a model, the internal reaction force of the material acting on the mass m depends not only on velocity and displacement but also depends on the rate of change of the relative acceleration, that is F \u00bc F\u00f0x; _x; &x\u00de (2) which is not commonly assumed. The study also presents suitable experimental research in this area conducted with a computer simulation technique on the selected sample a system with two and a half degrees-of-freedom. 1Corresponding author. Contributed by the Design Engineering Division of ASME for publication in the JOURNAL OF COMPUTATIONAL AND NONLINEAR DYNAMICS",
            " Manuscript received March 16, 2016; final manuscript received July 12, 2016; published online September 16, 2016. Assoc. Editor: Paramsothy Jayakumar. Journal of Computational and Nonlinear Dynamics JANUARY 2017, Vol. 12 / 014501-1 Copyright VC 2017 by ASME Downloaded From: http://computationalnonlinear.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jcnddm/935704/ on 02/02/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use The mathematical analysis involves determining the differential equations of the motion of the mass m in the system shown in Fig. 1(b), if the external excitations act only on the complex dynamic system (CDS) to which the mass m is attached to (no excitation forces directly acting on the mass m!). External excitations pl(t) may be in any form (continuous, discrete, determined, random, etc.). It is assumed that the variable xm describes the absolute displacement of the mass m and the variable xA the absolute displacement of the second end of the tested element fixed to the complex dynamic system at a certain location A (Fig. 2)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003220_fie.2016.7757483-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003220_fie.2016.7757483-Figure3-1.png",
        "caption": "Fig. 3. The robot wheel dimensions used by the waypoint project to calculate the rotations needed to navigate the course accurately.",
        "texts": [
            " In the second project, the goal is to build and demonstrate a waypoint navigator vehicle. When waving to the vehicle, the servo motors start to accelerate and navigate a predetermined course shown in Fig. 2 and return to its start location as soon as possible. The start line is marked by a foot-wide piece of black tape. The course is half foot wide marked with black tape. This project requires some algebra and rotational dynamics skills. Using the optical encoders and the wheels dimensions (shown in Fig. 3, the students can accurately and consistently navigate the robot through the waypoint. However, the challenge in this project is the open-loop system: the robots do not provide feedback for the user on their location. Therefore, the students have to carefully select their acceleration and speed to have enough friction on the wheels. The students were asked to form teams of 3 to 4 members. Each team was given a complete robot with additional electronics parts, like the ultrasonic sensors, touch sensors, resistors and wires"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001742_s00542-015-2460-4-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001742_s00542-015-2460-4-Figure2-1.png",
        "caption": "Fig. 2 The FEM model of HDD",
        "texts": [
            " In the transient shock simulation, a linear spring was used to model the air bearing between the slider and disk, and the calculated vertical force and moment between the slider and disk were applied to the slider dynamic simulation as input parameters. The slider dynamic simulation was able to describe the nonlinear air bearing system by considering the flying height of slider. The response of non-ramp\u2013 disk contact and contact was compared. 2.1 Transient shock simulation A finite element method (FEM) model of a 2.5-inch stamped HDD was constructed to model the structural dynamics using ANSYS Workbench, as shown in Fig. 2. The HDD model included the following components: the stamped base, cover, disk, motor, pivot, e-block, suspension and slider. A transient shock was used to obtain the vertical force and moment between the slider and the disk, which were extracted from the relative displacement between slider and disk. The slider and disk were connected using an air bearing spring, which was modeled as a linear spring even though the stiffness of the air bearing changes with the flying height of the slider. The feasibility of a linear spring air bearing model was investigated by Liu et al"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure1-1.png",
        "caption": "Fig. 1 Local 3D solid model of parts",
        "texts": [
            " The objectives of this task are to check if the gaps between blocks with blocks of IWS and blocks with VV supplied by designer can meet assembly requirements, and the sensitivity and contribution on different blocks/plates are calculated by worst-case searching and statistical analysis. This paper focuses on the tolerance analysis of IWS by tolerance analysis software CETOL inset of CATIA. The relationship between the sizes and locations of all solid parts is established in 3D space in CETOL platform. So the relationship between sizes and locations of all directions in all parts is fully considered. The analysis results are more accurate than the calculation of general plane dimensional chain [4]. As shown in Fig. 1, local 3D models nearby the equatorial port of ITER VV (seen from outer shell) were extracted, which included four block models of IWS and part of VV assembly models, and were assembled together. Each block in this field contains upper bracket, under bracket, ten plates, gaskets, washers, bolts and nuts, they are assembled together by two bolts. The under bracket with the upper bracket are connected by the middle of the bolt and nut, and are installed to VV ribs through the upper bracket and under bracket"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001325_iedec.2014.6784673-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001325_iedec.2014.6784673-Figure2-1.png",
        "caption": "Figure 2: Typical drawing practice",
        "texts": [
            " Fluid machines, namely, types, classifications, technical characteristics, rotor-performance. Pump performance, namely hydraulic heads, power and efficiency, performance curves, cavitation, performance testing. Centrifugal pumps, namely, functional principles, impellors, configurations, blade design and geometry, performance specifications and solid modeling. The computer software package CATIA [9] was used to teach the CAD section of the course, with exercises such as \u201cdesign of an angle bracket\u201d as shown on Figure 2, used for practice. The second section of the course was the Design Project, which lasted 8 weeks. The scenario for the project was that \u201cnew pumps were required for the cooling system of a large power station. To win this contract a more efficient version of the existing pump design is needed but there was insufficient resources to redesign and manufacture an entirely new pump. So the decision was taken to use an existing casing and redesign only the rotor and to save money, design and assessment will be performed on a small model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003236_detc2016-59619-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003236_detc2016-59619-Figure3-1.png",
        "caption": "FIGURE 3. A TYPICAL STATIC HEXAPOD MANIPULATION AND ITS COORDINATE FRAMES",
        "texts": [
            " This simple example is effective, because it may be formulated to directly produce the SESC model for the simple branched chain, primarily by following the computing direction. Kinematic analysis of the hexapod robot manipulation Hexapod robots have the highest efficiency for statically stable walking and can use one, two or three legs to function as hands [13,14]. For these reasons, hexapod robots exhibit superiority to other robots in terms of mobile manipulation. We analyze typical static hexapod manipulation by using adjacent legs in this paper, which can be modeled as in Figure 3. We define the frame \u03a3p as the universal coordinate frame attached to the ground, the frame \u03a3R as the body coordinate frame fixed on the body center, and \u03a3O as the object frame fixed on the object center. For each target leg, \u03b2i = [\u03b1i1,\u03b1i2,\u03b1i3] represents the joint angles; we attach a frame Si to the base of the leg and a frame Fi to the leg tip at the contact point. Note that the frame Fi moves with the leg tip, while the frame Ci, also located at the contact point, moves with the object. For each supporting leg,\u03b8 j = [\u03b1 j1,\u03b1 j2,\u03b1 j3] represents the joint angles, the framesB j are the base frames of leg j, frames Pjare located at the leg-end points and framesL j are at corresponding points on the ground (where,i= 1 \u00b7 \u00b7 \u00b72, j = 1 \u00b7 \u00b7 \u00b74)",
            " Then following the similar procedure depicted in Figure 2, we construct the SESC model for the manipulation. However, here we model each branch by using the SESC method, then combine the virtual COM of each branch to get the COM of the system by a classic method. To implement the real time control, a classic velocity based control method is adopted and the Jacobian matrix for the COM of system is provided. The position of the COM could be expressed as, x\u0304 = 1 mr +mo +\u2211 6 i=1 mi (mr x\u0304r +mox\u0304o + 6 \u2211 i=1 mix\u0304i) where, x\u0304i is the location of COM for leg i as shown in Figure 3, and it is modeled by the end-point position of a 3 DoFs of the SESC model. mi is the mass of leg i, mr and mo are the masses of the robot body and the manipulation object, and their COM positions are denoted as x\u0304r and x\u0304o separately. Then differentiating this equation with respect to time, we get \u02d9\u0304x = 1 mr +mo +\u2211 6 i=1 mi (mr \u02d9\u0304xr +mo \u02d9\u0304xo + 6 \u2211 i=1 mi \u02d9\u0304xi). (5) For a vector X\u0304 = [x,y,z] in 3D Euclidean space, the operator \u2227 is defined as, \u02c6\u0304X =  0 \u2212z y z 0 \u2212x \u2212y x 0  Then the linear velocity of each contribution to the COM can be derived as \u02d9\u0304xr = [ I3\u00d73 \u02c6\u0304xPR ]T AdgPRVPR = [ I3\u00d73 \u02c6\u0304xPR ]T AdgPRJ+X Ja\u03b8\u0307a, (6) \u02d9\u0304xo = [ I3\u00d73 \u02c6\u0304xPO ]T AdgPOVPO = [ I3\u00d73 \u02c6\u0304xPO ]T AdgPO [G1,G2] \u2212T A[\u03b8\u0307a, \u03b2\u03071, \u03b2\u03072] T (7) and \u02d9\u0304xi = [ I3\u00d73 \u02c6\u0304xPxi ]T AdgPxi VPxi = [ I3\u00d73 \u02c6\u0304xPxi ]T AdgPRJ+X Ja\u03b8\u0307a+ [ I3\u00d73 \u02c6\u0304xPxi ]T AdgPxi J \u2032b bi \u03b8\u0307i, (8) 4 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002312_s0219519416500998-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002312_s0219519416500998-Figure1-1.png",
        "caption": "Fig. 1. Schematic of various positions measured at intervals with a high-speed camera. Positions at the commencement of a sprint were recorded at 4 main intervals: (a) setting off position, (b) push-off, (c) first step landing, and (d) second step landing.",
        "texts": [
            " Equipment A high-speed camera (Inline, Fastec Imaging Corp., USA) with a sampling rate of 250Hz and a shutter speed of 1000Hz was used to collect motion images in the sagittal plane. The area used to capture video was set up with a width of 5m and the height and distance of the participants relative to the camera were 1.2m and 10m, respectively. In addition, the camera was set perpendicular to the direction of motion. Images were obtained for setting off, pushing off, and the landing of the first and second steps (Fig. 1). The image resolution was 680 480 dpi. An International Association of Athletics Federations approved push-off block (F155A, Nishi Ltd., Japan) was used in this series of experiments. 2.3. Procedures Participants were asked to warm up for 20\u201330min prior to the commencement of testing. After warming up, participants were asked to run as fast as they can until 1650099-3 J. M ec h. M ed . B io l. D ow nl oa de d fr om w w w .w or ld sc ie nt if ic .c om by M O N A SH U N IV E R SI T Y o n 04 /1 0/ 16 ",
            " In the sprinting process, the first discriminated movement in sprint start after gunshot was the head lift up. Thus, the beginning of the sprint start for data analysis was indicated by a head lift. The push-off was defined as the commencement of the sprint to block clearance and defined as the point that both the front and RF left the block. The kinematic parameters measured included the COM resultant and vertical velocity, the peak foot linear velocity, height of COM, trunk angle, take-off angle (Fig. 1(b)), and step length (distance from front to rear toes). The trunk was defined as a vector from hip to shoulder, and the trunk angle was the vector relative to the horizontal (Fig. 1(a)). The trunk vector was defined as the minus angle when the vector is under the x-axis in the coordinate system. 1650099-5 J. M ec h. M ed . B io l. D ow nl oa de d fr om w w w .w or ld sc ie nt if ic .c om by M O N A SH U N IV E R SI T Y o n 04 /1 0/ 16 . F or p er so na l u se o nl y. 2.5. Statistical analysis SPSS (Version 20.0, SPSS Inc., USA) was used to compare and analyze all kinematic data. Statistical significance was set at \u00bc 0:05 for all tests. A one-way analysis of variance (ANOVA) with repeated measures was performed to determine the kinematic differences in three different crouched starting positions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002778_j.proeng.2016.07.095-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002778_j.proeng.2016.07.095-Figure4-1.png",
        "caption": "Fig. 4.Construction of a tooth rim of the gear: (a) involutes equation; (b) tooth contour; (c) gear tooth pressed out along the helix; (d) 3D model of the gear",
        "texts": [
            " 2): by the command Plane we specify over a distance of 50 mm from the plane of the right view a plane, parallel to it and a face of the gear blank; in the created plane we make a sketch of the circle of the helix base with a diameter, which is equal to the reference diameter, and by the command Helix and Spiral we set the parameters: height 100; constant pitch 2896,2225; clockwise; initial angle6,215\u00b0. The initial angle which sets distribution of the initial point O of the We make a base circle dvb1of the equivalent wheel (Fig. 4, b). By the command Equation Driven Curve, having chosen the radius of the equivalent wheel dvb1/2 = 80.88735 (Fig. 4, a), we make an involute to a circle e. Moving back from the point P for mn = 4 mm and 1,25mn = 5 mm, we draw addendum and dedendum circles limiting the tooth contour. We form a constant chord of the tooth sc = 5,55mm at a distance hc = 2,99 from the tooth top (see Fig. 4, b). Using rotation around the center Cv and the command Mirror Entities, we create from the involute ear closed tooth contour, which goes through limiting points of the segment sc. By the dedendum circles using the command Extruded Cut we cut the exteriors of the blank. We extend the tooth contour by the command Swept Boss/Base, having set the parameter Merge Result and indicated the helix We cut the tooth, which goes beyond the limits of the gear, with the help of the command Extruded Cut, having performed the fillet curve of the tooth f = 0,4mn = 1,6 mm; the chamfer 2 \u00d7 45\u00b0 (Fig. 4, c). By the command Circular Sketch Pattern we distribute 41 teeth over the surface of the root cylinder, by so doing we finish construction of 3D model of the gear (Fig. 4, d). To construct the wheel rim we need to repeat the same operations as for the gear, but with the parameters for the wheel. We open a new document 3D arrangement of parts and/or other assemblies. We activate the command Layout at the panel Layout and make a layout sketch (Fig. 5, a). We specify the pitch point P in the initial point and draw from it two vertical segments, which are equal to the reference diameters of the gear and the wheel. Though the bisecting point of the segment we draw the gear and wheel axes, the length of which is equal to their width"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002196_1468087414556134-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002196_1468087414556134-Figure1-1.png",
        "caption": "Figure 1. Asperity contact between two rough surfaces.",
        "texts": [
            " The models of the piston ring friction developed in this study include a mixed lubrication model considering the full/starved oil supply and a dry running friction model. Meanwhile, an elasto-plastic contact model is also added in the piston ring friction model. The mixed lubrication model consists of hydrodynamic lubrication model and asperity contact model. The formulations and boundary conditions of the mixed lubrication model can be found in Guo et al.28 Instead of asperity contact model in the conventional mixed lubrication model, an elasto-plastic contact model was employed in this study. As shown in Figure 1, if the value of oil film thickness is small, the asperity contact may occur. The assumption of elastic contact is suitable for the small deformation, but the asperity contacts may contain large deformations under heavy load and low sliding velocity conditions, which makes the above assumption unavailable. In this study, the asperity contact model should be divided into elastic asperity and elasto-plastic asperity contact models. As shown in Figure 1, when the asperities on the two rough surfaces contact, the interference v at the contact is defined as v= z1 + z2 h 2f r 2 \u00f01\u00de 2f r 2 = r2 4R \u00f02\u00de where h is the nominal clearance between two contacting surfaces, R is the radius of curvature at the asperity peak, z1 and z2 are the heights of asperities from the mean of asperity heights and r is the factor considering the misalignments between asperities of two rough surfaces.6 A critical interference derived from the von Mises yield criterion is defined to distinguish between the two contact models29 vc = pDSy 2E0 2 R \u00f03\u00de where Sy is the yield strength, E0 is the composite elastic modulus of the contact surface materials, R is the radius of hemispherical asperity and E0 and D can be defined as29,30 E0= 1 1 v2 1 E1 + 1 v2 2 E2 h i \u00f04\u00de D=1:295 exp (0:736n) \u00f05\u00de where E1, n1 and E2, n2 are the elastic modulus and Poisson\u2019s ratio of the ring and cylinder liner, respectively, and n is Poisson\u2019s ratio of the material which yields first"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002943_s13369-016-2309-x-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002943_s13369-016-2309-x-Figure1-1.png",
        "caption": "Fig. 1 Robotic coax-helicopter prototype",
        "texts": [
            " Theory 2 There is a stabilizing controller K , if and only if there exist positive definitive matrices R and S such that: \u23a1 \u23a3 AR + RAT \u2212 \u03b3 \u03bb\u22121BBT RCT \u2212 L CR \u2212 \u03b3 I I \u2212LT I \u2212 \u03b3 I \u23a4 \u23a6 < 0 (9) ATS + SA + CTLT + SLC \u2212 \u03b3CTC < 0 (10)[ R I I S ] \u2265 0 (11) These inequality matrices provide a larger feasible region of solution over the Theory 1. However, it has to solve two Riccati type inequalities with a matrix equation. To avoid this disadvantage, simple but sufficient linear conditions is proposed for the parametric H\u221e loop shaping based on LMI in the following section. A robotic coax-helicopter prototype was built showed in Fig. 1. The basic rigid body of prototype focuses on the origin of mass. Like researches of control design for most helicopters, the basic 6-DOFs rigid body describes theirmain feature bypassing the extra degrees, such as coning and tilting freedom introduced by blades being loosely connected with fuselage [27\u201329]. At the same time, a dynamic model with 6-DOFs rigid body and generation mechanism of aerodynamic forces and torques could satisfy the purposes of control design through a Bell 205 flight test [27]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001469_isie.2014.6864938-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001469_isie.2014.6864938-Figure1-1.png",
        "caption": "Fig. 1. Seven phase half-coiled winding configuration",
        "texts": [
            " Then, the analysis formulation of the inductance and suspension force are deduced and verified by Finite Element Analysis (FEA). Furthermore, the control system is proposed based on the 978-1-4799-2399-1/14/$31.00 \u00a92014 IEEE 2080 derived voltage model, torque model and suspension force model with the rotor eccentricity being considered. Finally, the Simulink simulation is provided on a prototype motor to verify the validity of the analysis. II. WINDING CONFIGURATION AND HARMONIC ANALYSIS OF SEVEN-PHASE BEARINGLESS MOTOR A seven-phase 28-slot bearingless motor with half-coiled winding is illustrated in Fig.1. Each phase winding is composed of 4 coils. The axes of the seven phase windings are 2/7 apart in the space. All the phase windings are star connected. The MMF distribution for the seven-phase half-coiled winding is shown in Table I based on the analysis method presented in [22], in which the relative amplitude of the MMF is given numerically. And the \u201cF\u201d sign indicates those MMFs which are rotating forward, while the \u201cB\u201d sign represents those MMFs rotating backward. N is the number of series turns per phase, I is the magnitude of phase current, y1 is the coil pitch, v is the space harmonic order, u is the time harmonic order and  is the pole pitch"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure11-1.png",
        "caption": "Fig. 11. Simulation of passing through critical clearances",
        "texts": [
            " Considering the lowest value, mechanical safety factor can be determined as 2.5. If directions of forces and torques are also taken into account, this safety factor increases to nearly 3. V. ELECTRICAL SIMULATIONS Besides the mechanical analysis of the structure, evaluation of electrical aspects is also particularly important. For the inspection of electrical stresses, potential and electric field distribution has been examined in case of a critical geometry, where the clearances are close to the minimal distance required for safe live-line work in Hungary [2]. Fig. 11 shows a grounded tower structure and the phase conductors in the middle phase held by double composite insulators: it is a typical arrangement in the Hungarian 400 kV high voltage grid. Fig. 12 shows a closer picture of the simulated arrangement focusing on the model of the human body and the new conductor car. Human body model was created based on the guidance of IEC 62233 [3] with some modification required because of the sitting position of the worker during the simulations. In this case finite element solution of COMSOL MultiPhysics was used to determine both the electric potential and the electric field distribution in the vicinity of the arrangement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000877_cict.2015.159-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000877_cict.2015.159-Figure4-1.png",
        "caption": "Figure 4. The proposed hexarotor body with payload",
        "texts": [
            " III. MOTOR EVALUATIONS Motor and propeller selection is the next step of the product design. We consider outer rotor type brushless DC motors for our product. This is due to availability matter and weight of motor and the amount of thrust could be gained from each motor. The following is comparison of thrust test results between two outer rotors. Power thresholds in the figure indicates the peak per converter. We selected motor gives more than 10kg about 80% throttle with 29*9.5CF propellers. Figure 4 illustrates the proposed hexarotor body with 30kg payload. To achieve aerodynamically optimum the body is designed to be compact. In this last section of design we proposed a unique cable winding device. Since the power supply to aerial vehicle is 380V DC through cable during this research, it is necessary to consider the cable winding carefully to avoid excessive tension on power and signal supply cable to prevent hazards. To accomplish this task we designed an automated cable winding structure that wind and rewind the supply cable automatically or manually"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001521_cphc.201301101-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001521_cphc.201301101-Figure1-1.png",
        "caption": "Figure 1. a) Top: Ellipsoid on a cone presenting a molecule tilted and oriented as given by order parameters. a) Bottom: Bird\u2019s eye perspective on the cone and the simplified presentation of the layer structure with the tilt and the polarization order marked as arrows. b) Top: Bent-shape molecule with the same tilt presented in two possible favourable orientations\u2014gray and white. b) Bottom: Symbolic presentation of the molecular side view for both orientations presented above. The bold line gives the tilt of the bent molecules. The arrow notation shows polarization pointing toward or away from the reader.",
        "texts": [
            " Molecules forming antiferroelectric phases have two CO groups, resulting in a larger molecular dipole. The arrangement of more dipoles often contributes to a significant molecular quadrupole as well. Finally, all molecules have chiral groups attached at the end of the molecular core influencing rotation around long molecular axes. Although structures of phases found in antiferroelectric liquid crystals are known to researchers studying them explicitly, it is worth to describe them in more detail. Schematic representations of the structures are presented in Figure 1 a. Molecules are organized in smectic layers and they are tilted away from a layer normal. The magnitude of the tilt is described by an angle q that the average long molecular axis forms with the layer normal, usually corresponding to the z coordinate axis. Another important piece of information for the structure is the tilt direction that is given by an angle f measured as an angle between the tilt projection onto the smectic layer and the chosen direction within a smectic layer corresponding to x coordinate axis. As molecules in the layer may tilt in any average direction, the structure of a single layer is presented as a cone with an apex angle 2 q and the average direction within the layer is shown as an ellipsoid on this cone (Figure 1 a). Such presentations are often stylized further for more complex structures. The cone representing a layer is given as a circle in a bird\u2019s eye view and the arrow gives the tilt direction (Figure 1 a below). For more complex structures with modulations over several layers, the projection of cones (circles) on one smectic plane is given. In order to follow the structure from layer to layer, arrows denoting tilt directions are marked with Table 1. Standard materials exhibiting phases found in antiferroelectric liquid crystals. The phase sequence upon lowering temperature, starting from orthogonal SmA phase, including all existing tilted polar phases, is given for optically pure materials with an exception of MHPOBC, that is slightly racemized and has more phases in a sequence then a pure one",
            " Due to the very short pitch of the modulation the structure of the SmC* a phase is optically similar to the SmA phase. For the SmC* Fi1 and the SmC* Fi2, the resonant X-ray scattering has shown strictly commensurate periodicities. The modulation of the SmC* Fi1 phase extends over three smectic layers. The direction of tilts in neighboring layers differs for either phase difference a or for the phase difference b, which are not equal. The sequence of phase differences is well defined as phase difference a is always followed twice by phase difference b (see Figure 1 e). The commensurate period of the elementary unit exists because the sum a+ 2 b 2 p. A slight deviation of the sum from 2 p results in an additional helical modulation on the scale of several hundreds of smectic layers. The modulation of the SmC* Fi2 phase extends over four smectic layers. Its structure is also defined by a sequence of two different phase differences a and b that interchange. The sum a+ b p is shown in Figure 1 d. The slight deviation of the sum of both phase differences from p, again results in a helical modulation on much longer scale. Finally, the elementary modulation of the most recently discovered SmC* 6d phase extends over six layers. It appears strictly below the SmC* a phase. The structure is very similar to the SmC* Fi1 phase as two different phase difference interchange in the sequence a,b,b, but the sum of the three angles in the sequences is a+ 2 b p (Figure 2 b). In 1995 new systems of polar smectics formed of molecules with a bent core were discovered",
            "[12] Phases were named according to the time of their naming as Bx, for example B1, B2 and so on. In this concept paper we discuss only structures where bent-shaped molecules organize in smectic layers that are polarly ordered within the B2 phase. Within this phase the whole set of structures is stable and the structures were given descriptive names. Several types of structures within B2 are possible.[13] To describe those structures let us first introduce the graphical presentation of order in bent-core systems. As seen in Figure 1 b (top right), the polarization is associated with the bent orientation and is marked as an arrow. The tilt is associated with the orientation of the line that schematically connects ends of the core. As the polarization can have two directions with respect to the tilt, it is necessary to present the polarization and the tilt in each layer. The tilt is marked as a bold line and the polarization is marked with arrows, as described in details in the caption of the Figure 1 b. We discuss six types of polar structures that are considered as subphases of the B2 phase. The bent molecules organize in a layered order bent in one direction only. As CO groups are found close to the middle of the molecule, the average orientation of the bent defines also the direction of the polarization. As the polarization can be influenced by an external field, the polarization is used for a description of the bent orientation. Two structures have the order equal in all layers. In the recently discovered SmAPF structure phase molecules in layers are not tilted and polarizations have the same direction (ferroelectric order) in all layers",
            " Let us shortly present the most elaborate free energies for the two systems described in introduction for the chiral polar smectics formed of elongated chiral molecules and for the achiral polar smectics where polar ordering is induced by ordering of the bent-shaped molecules in the layer. Free energies presented in continuation allow for all experimentally confirmed phase structures and phase sequences.[17, 18] For description of the layer order we assume that the nematic and the smectic orders are constant and we further limit our studies to two vectorial order parameters\u2014the tilt and the polarization, which are defined in the same way for elongated as well as for bent-shaped molecules (Figure 1). The tilt order parameter for molecules in the jth smectic layer ~xj recapitulates the quadrupolar nature of the up-down molecular packing symmetry [Eq. (1)]: ~xj \u00bc fnj;x nj;z; nj;y nj;zg \u00f01\u00de where ~nj \u00bc fnj;x; nj;y; nj;zg is a director in the jth smectic layer. The director is related to the direction of average long molecular axis in chiral polar smectics (Figure 1 a) or to the direction of the longest dimension of the bent-shaped molecule (Figure 1 b). The polarization in the chiral smectic liquid crystals is an improper order parameter and it is induced by the tilt. When a chiral molecule is tilted and encircled by other tilted chiral molecules, rotation around long molecular axis is hindered and one favorable orientation exists due to the chiral symmetry of the molecule. The molecule spends more time in this position, the molecular dipoles do not cancel out on average and the layer is polar. An isolated tilted layer is polarized perpendicularly to the tilt due to symmetry reasons,[1] but in more complex tilted structures the polarization can have a general direction.[19] In fact, the hindrance of the rotation is a steric effect and the polarization is the consequence only. Its direction is parallel to the average molecular geometric polar axis which is not necessary associated to the direction of the molecular dipole. However, the average polarization has the same direction as an average geometric axis, it is a property that is measured easily and we suggest that although by origin geometrical, the order parameter is called polarization (Figure 1 a), given by Equation (2): ~Pj \u00bc fPj;x; Pj;yg \u00f02\u00de The reasoning described above is even more evident in systems of bent-shaped molecules. Imagine rotation of the bentshaped molecule encircled by other bent-shaped molecules. As a molecule rotates around the long molecular axis, two orientations are more favorable than others. Out of the two orientations, given as a grey and a white molecule in Figure 1 b, one of them is the most favorable, because then molecules can pack the most tightly. Therefore, ordering of bents is preferred sterically. Molecular dipoles due to the CO groups close to the central benzene ring is parallel or antiparallel to the 2014 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim ChemPhysChem 0000, 00, 1 \u2013 14 &5& These are not the final page numbers! second molecular axis defined by the bent order, and the layer with ordered bents is polar even without a tilt. As the polarization of systems formed of ordered bent shaped molecules may be polar even without a tilt the polarization is a proper order parameter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001149_2014-01-1797-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001149_2014-01-1797-Figure16-1.png",
        "caption": "Figure 16. Architecture of the motor",
        "texts": [
            " In order to use motor torque effectively to advance the performance of the second generation of the system, we undertook the challenge of reducing the margin and adopted a control method for the system that regards the difference between the command value and the actual clutch torque as an external disturbance and reduces the difference through feedback control. Owing to this integrated control of the motor and clutches, there was no need to set the command values of the system higher, making it possible to increase the motor torque for both engine start and EV drive. Figure 16 shows the architecture of the motor. Increasing the motor torque is important for improving both fuel economy and acceleration performance. In order to leverage the full potential of the motor and increase the motor torque without changing the size, weight and materials, it is effective to clarify the relationship between the input load and the durability line of the motor. Figure 17 shows the flowchart used for confirming component durability. We applied this procedure to the motor. First, we defined the stator coil films as being the weakest part with respect to service life, and the coil film temperature and output time as their input load factors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001154_j.electacta.2014.12.151-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001154_j.electacta.2014.12.151-Figure1-1.png",
        "caption": "Fig. 1. Scheme of the flowing electrolytic device.",
        "texts": [
            " The special designed flowing electrolytic device has a large electrode area, and the flowing deionized water which was used as electrolyte kept washing the electrode surface. Two kinds of carbon nanomaterials with different morphology such as CNPs and carbon nanosheet(CNSs) were generated and separated in the [(Fig._2)TD$FIG] flowing electrolytic device, and their properties and potential applications were also discussed. The electrolysis was occurred in a homemade flowing electrolytic device as shown in Fig. 1. The anode of the flowing electrolytic device was a graphite cylinder. Its diameter and height were 18mm and 100mm, and its working areawas 5600mm2. The large anode area was beneficial to generate more products and convenient for further enlargement. The graphite anode was placed in the center of a titanium tube cathode. The inner diameter and height were 20mm and 102mm. Hence, the gap between the anode and the cathode was only 1mm. The narrow gap leads to a high intensity electric field. The high electric intensity enabled that the electrolysis occurred in deionized water"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.24-1.png",
        "caption": "FIGURE 8.24",
        "texts": [
            " These results indicate the accuracy of the model translation and reassembly and imply that the quarter suspension kinematic analysis can be duplicated in SolidWorks. Result graphs: (a) shock travel and (b) camber angle. Verification of kinematic analysis results: (a) SolidWorks vertical wheel travel; (b) SolidWorks shock travel; (c) SolidWorks camber angle; (d) Pro/ENGINEER vertical wheel travel; (e) Pro/ENGINEER shock travel; (f) Pro/ENGINEER camber angle. The dynamic analysis of the quarter suspension was performed by taking racecar weight, spring rate, and shock damping into consideration. As shown in Figure 8.24, a 150 lb external force pointing upward was applied on the road profile cam to mimic the wheel load due to racecar weight (445 lb) and driver weight (155 lb). An equilibrium analysis was first carried out. The equilibrium state of the racecar was assumed as the initial condition for the dynamic simulation, in which the racecar started in equilibrium on the flat road and then reached the first hump. A spring and a damper were also defined in the dynamic analysis, as shown in Figure 8.25. The physical position of the spring is shown in Figure 8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002944_icma.2016.7558946-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002944_icma.2016.7558946-Figure3-1.png",
        "caption": "Fig. 3. Deformation in the effect of gravity. Vascular model is placed on a plane (left). The nodes on one side of the vascular are rigidly fixed (right).",
        "texts": [
            " The less the number of triangular patches, the faster calculation speed, but the worse resolution of the model. So, the calculation speed and models\u2019 resolution must be take into account simultaneously. The strategy of geometry model closely envelope the topology network is used to construct the vascular model, as shown in Fig. 2. In this figure, each point in topology network represents a mass element, and each line represents a damping spring connecting two mass elements. The deformation of the vascular model in the gravity field is shown in Fig. 3. In Fig. 3a, the vascular model\u2019s cross section deformed in an elliptical shape under the effects of gravity, and the model has an obvious refractive effect, as shown in Fig. 3b. The biomechanics properties of biological vascular substitutes(BVS) in Francois Auger\u2019s laboratory were tested by Georgia Institute of Technology [8]. The biomechanics properties of axial tension and radial expansion of vascular model is analyzed, and the experimental results are compared with the results obtained by Georgia Institute of Technology. The mechanical properties of biological soft tissues are characterized by the viscoelasticity, anisotropy, stress relaxation and creep [9], [10]. When the soft tissues suffer cyclic load, the unloading curve normally falls below the loading curve to form a hysteresis loop"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002697_acc.2016.7525278-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002697_acc.2016.7525278-Figure1-1.png",
        "caption": "Fig. 1: Sketch of the EPS system.",
        "texts": [
            " (11) The objective of the subsequent section is the application of the proposed concepts to an electric power steering system in order to recover states as well as unknown inputs. With regards to the modelling of the EPS system the state variables commonly refer to both angular velocities and positions, e.g. [10]. However, for the actual application, i.e. recovery of the unknown inputs driver steering torque and exogenous wheel torque, a model as discussed in [12] suits the requirements ideally. A sketch of the simplified EPS system structure is illustrated in Fig. 1. Basically, its main components are the steering wheel and column, the servo unit (electric motor) assisting the driver, and the steering rack that is connected to the column via a pinion. The steering angle \u03b4h and its velocity \u03c9h := d\u03b4h dt are both assumed known. Whereas the former is measured, the latter can be calculated straightforwardly by a robust exact differentiator [21]. Other than that, a torque sensor is located right above the pinion (from the perspective of the steering wheel) providing the column torque"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000362_8611_2011_57-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000362_8611_2011_57-Figure15-1.png",
        "caption": "Fig. 15 Schematic view of TWB joints",
        "texts": [
            " There are numerous and diverse applications of laser welding into the automotive industry, being the main ones: \u2022 Multi-thickness welded blanks; \u2013 Tailored welded blanks: Blanks of relatively simple geometry with linear weld seams or welded blanks of complex shape with non-linear weld seams, for high productivity, weight optimization and resistance improvements; \u2013 Welded patchwork blanks for components requiring local reinforcement. \u2022 Car assembly parts, \u2022 Body in white elements, \u2022 Mechanical elements. One of the main changes in the automobile manufacturing process was reached with the introduction of tailor welded blanks (TWB) for car parts manufacturing. Tailor blanks are composed of two or more dissimilar sheet metals, with different thicknesses, shapes, strengths, or materials that are butt-welded together before being formed. Figure 15 shows the joint configuration of a tailor welded blank. Initially, TWBs were used mainly to reduce material consumption and cost. Thus, for example, pieces of stamping scrap can be jointed together, stamped again and reused as a new part for reducing material consumption. The classic manufacturing sequence, i.e., first form and then join, is inverted by this innovative product. While in the conventional parts manufacturing process, two or more stamped parts are spot-welded together to form a part, in the TWB stamping process, sheets are first welded together, and then integrally stamped into a part"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001008_recl.19590781110-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001008_recl.19590781110-Figure3-1.png",
        "caption": "Fig. 3. Polagrams of manganate on 0.1 M NaOH.",
        "texts": [],
        "surrounding_texts": [
            "I . M. Issa, R. M. Issa and I . F. Hewaidy, - ~ ~ ____ 892 -\nas the alkali concentration is increased to 1 M , the diffusion current of the first wave decreases to a greater extent than that of the second.\nBetween 1 and 2 M NaOH the decrease in the diffusion current becomes smaller than the decrease occurring between 0.5 and 1 M NaOH. Another fact which can be deduced from the curves is that the second wave is better developed at lower than at higher alkalinities. These results are in general in conformity with the findings of den Boef and Poeder 5. However, these authors could not obtain a well defined wave for the Mn+*/Mno step.\nVariation of the diffusion current with manganate concentration On plotting the diffusion current measured at three different values of applied potential viz. at -0.3 V, -1.5 V and -1.85 V at two different alkalinities, the curves shown in Figs. 3 and 4 are obtained. The curves indicate that in 0.1 M alkali, the limiting current is pro-",
            "Studies in polarography IV. 78 (1959) RECUEIL 893 - ~~ ... - .. ~~ -\nportional to the manganate concentration at all three potentials within the concentration range 0.25-1.25 millimolar. In 0.5 M NaOH the concentration range within which a linear relationship is obtained extends up to 2 millimolar. Den Boef and Poeder 5 also found a linear relationship between the diffusion current measured at -0.4 and -1.4 V vs S.C.E. and the manganate concentration from 2 X 10-4- 2 X lo-* A4 at the D.M.E.\nEffect of Formic acid The polarograms shown in Fig. 5 indicate that when 1, 1.5 and 2.5\nequivalents of formic acid are used in the reduction process, (i.e. in the presence of 0, 0.5 and 1.5 moles in excess of those necessary for the reduction of Mn0,- to Mn0,-2), the polarograms recorded soon after mixing were almost the same. However, on allowing the manganate-formate mixture to stand for 15 minutes, the diffusion current decreased 2.7 \"/o and 15.3 % in the presence of an excess of 0.5 and 1.5 moles of formic acid respectively. No decrease in the diffusion current was obtained when no excess formic acid was added.\nD i s c u s s i o n The usual manner for computing the mode of reduction a t the drop-\nping electrode is to calculate from the diffusion current the number ._ -\nJ . Heyrovsky and D. Ilkovic, Collection Czechoslov, Chem. Communs 6, 498 (1934).",
            "894 I . M. Issa, R. M. Issa and I . F. Hewaidy,\nof electrons involved in a given process. This is done by the aid of the Ilkovic-Heyroosky equation 4 from a knowledge of the capillary characteristics m and t and the diffusion coefficient D of the electroactive species. This is not, however, possible in the present case owing to our ignorance of the diffusion coefficient of the manganate ion. Recourse was therefore taken to a different way of calculation depending upon the tacit assumption that reduction of manganate passes at -1.85 V to MnO, in the same way as the permanganate ion. The reduction of the manganate ion to this definite stage takes place according to the equation: M n 0 4 - ~ + 4 H 2 0 + 6 e + M n o + 8 0 H - . . . ( 1 ) By dividing the total current (at -1.85 V) by 6, one obtains a value if id corresponding to a single electron (x) in the reduction process. The number of electrons involved at the other stages can be calculated by dividing the diffusion current by the value of x (cf. Table I) . Since the diffusion current changes with the alkalinity of the solution, it is deemed necessary to calculate the value of (x) at each alkali concentration.\nT a b l e I\nValues of id per electron at different alkalinities\n::$ 1 {.I6 'A 1 :::; 'A 1 !z 0.75 1 .o 4.03 ,. 3.60 1.25 4 ,. 3.62 ,, - 1 S O 3.67 ,, 3.55 1 , 1.8\nConcentration x for x for Concentration x for 1 M m M NaOH NaOH NaOH of MnO,-e 1 0.1 M 1 0.5 M 1 of M 2 - 2\ni:: 'A 3.61 3.4 ,, -\nUsing these values, the number of electrons consumed at the different reduction stages in 0.1, 0.5 and 1 M NaOH were calculated as shown in table 11.\nThe potentials of -1.85 and -1.5 V were chosen for these calculations because there the diffusion currents are well defined. On studying the effect of mercury height ( H ) on the limiting current at different potentials in 0.5 and 1 M NaOH, it was found that at potentials less negative than -0.8 V the current did not vary appreciably on changing the mercury height from 46 to 83 cm. At -1.5 V and -1.85. V or higher, significant changes in the diffusion current occurred"
        ]
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure3-1.png",
        "caption": "Fig. 3 Frozen domain.",
        "texts": [
            " (2) instead provides the relation between the structural stiffness and the design variables given by the SIMP interpolation scheme. Starting from an initial solution (uniformly distributed material over the design domain), material distribution is iteratively updated until convergence to a minimum is reached. A finite element model of the brake caliper and the upright has been developed for topology optimization. The design and frozen domains referring to the caliper and the upright have been defined. The design domains of the brake caliper and the upright are depicted in Fig. 2. The frozen domain is shown in Fig. 3. The frozen domain is made by the hub bearing houses, the connections of the upright to the suspension wishbones and steering rod, by the connections between the caliper and the upright and by the six cylinders and pad supports in the brake caliper. This domain is fixed and is not involved in the optimization process. Design and frozen domains have been discretized with linear tetrahedral elements, the average mesh size was 2 mm. Aluminum alloy has been considered as reference material for both components"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002495_s1068798x16010159-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002495_s1068798x16010159-Figure1-1.png",
        "caption": "Fig. 1. Loading of a threaded joint by a tipping torque cre ated by a tensile (a) or compressive force (b) displaced from the symmetry axis; by a tipping torque (c); and by a tipping torque with rotation of the joint (d).",
        "texts": [
            "6 In experiments regarding the influence of the tip ping torque on the operation of a group threaded joint, the joint is loaded, as a rule, by tensile (Fig. 1a) [1, 2] or compressive (Fig. 1b) [3] force F displaced eccen trically with respect to the symmetry axis; or by a tip ping (bending) torque (Fig. 1c) [4, 5] created by two forces positioned symmetrically relative to the beam bearings. No experiments have been conducted to study a threaded joint under the action of a tipping torque with rotation of the joint (Fig. 1d)\u2014for exam ple, the threaded joint between the rim of a spiral gear and a disk gear\u2014evidently because of the difficulty of obtaining readings in this configuration with measur ing instruments mounted at the joint. The skew axis of a threaded joint is the axis where the pressure in the contact zone created by the tipping torque is zero [6]. The position of this axis determines the load on the screw created by the tipping torque. Experiments show that it is displaced relative to the neutral (geometric) axis of the contact plane toward the compressed side of the joint [1\u20135]",
            "\u2013= system are symmetric with respect to the support; the distance between the points is 430 mm. To measure the contact displacement close to the upper and lower edges of the threaded joint, two indicators (scale divi sion 0.001 mm) are mounted at a distance l = 47 mm from the neutral axis. The load is applied by means of a manual press, through a dynamometer. The dyna mometer reading of 2000 N corresponds to loading of the threaded joint by a tipping torque of 225 N m. We consider two series of experiments. (1) Tests corresponding to Fig. 1c. We conduct five loading cycles. A cycle includes resetting the indica tors to zero; applying a load of 1 kN; recording the indicator readings; applying a load of 2 kN; recording the indicator readings; removing the load; recording the indicator readings. If the readings do not return to zero, they are recorded and then reset to zero. The results of the first six cycles are shown in Fig. 6a. (2) Tests corresponding to Fig. 1d. We conduct eight loading cycles. A cycle includes resetting the indicators to zero; applying a load of 2 kN; recording the indicator readings; removing the load; recording the indicator readings; rotating the apparatus by 180\u00b0 around a horizontal axis. If the readings do not return to zero, they are recorded and then reset to zero. The results of the first six cycles are shown in Fig. 6b. Measurements of the roughness Ra at different points of the contact surfaces after the tests show that Ra1 = Ra2 = 11 \u03bcm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001089_icarsc.2014.6849805-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001089_icarsc.2014.6849805-Figure7-1.png",
        "caption": "Fig. 7. Hokuyo URG-04LX-GU01 LASER rangefinder sensor [15].",
        "texts": [
            " It is not necessary to convert the rotational speed because this is the same in the inertial frame.  [ \u0307 \u0307 \u0307] [ ] [ \u0307  \u0307  \u0307 ] [ [   C. Perception and location After obtaining the velocities of the robot in the inertial frame, it is necessary to determine its position on a theoretical frame, and determine also the obstacles. The robot position is obtained through integration of its velocities previously determined, according to (10).  [ ] [ \u222b  \u0307 \u222b  \u0307 ]   KUKA sells a youBot model with the Hokuyo URG-04LXGU01 LASER rangefinder sensor [15] (Fig. 7). The model of this sensor was attached in the front of the simulated youBot, to perform the perception of the surrounding environment. This sensor sends to the youBot script an array with the coordinates of the obstacles detected, in a range of 240\u00ba, in relation to the sensor reference. To convert these values to the robot reference (geometric center of the platform) a conversion is required according to (11) and (12).       being SLn the distance of the laser to the obstacle along the xaxis, and CLn the distance along the y-axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002232_icma.2014.6885882-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002232_icma.2014.6885882-Figure3-1.png",
        "caption": "Fig. 3. The D-H coordinate systems of laparoscope arm.",
        "texts": [
            " According to the structural features of general mechanical arm, the kinematics characteristics of the trocar points(the fixed points) of instrument arms and laparoscope arm are the same, and the passive joints determine the position vector of trocar and orientation of active joints\u2019 end effectors. In this paper, the DoFs of micro device are very closed and they influence the end\u2019s workspace less, so the passive joints\u2019 kinematics and instrument arms\u2019 workspace can be overall considered for convenient purpose, and these two problems are integrated in same D-H coordinate systems, as shown in Fig. 3, frame 0 0 0x y z , 1 1 1x y z , 2 2 2x y z and 3 3 3x y z are the coordinate systems of passive joints; frame 4 4 4x y z , 5 5 5x y z , 6 6 6x y z and 7 7 7x y z are the coordinate systems of active joints; frame 0 0 0x y z is also the base coordinate system, frame 8 8 8x y z is the tools coordinate system. The link parameters are shown in Table I. If we know the preoperative positioning parameters of passive joints, the surgical assistant can placed the 4 passive joints in the optimal position and they will be fixed by electromagnetic clutches, and the system collected the data of each joint by its encoder for basic information inputting to the kinematics algorithm. We can calculate out the transformation matrixes of near two links with the information from Fig.3 and Table I; through using the transformation function as show in equation (1), we can calculate out the transformation matrixes 0 1 2 3 4 5 6 7 1 2 3 4 5 6 7 8 T T T T T T T T . 1 , , , , C C C S S C S S S C S S 0 S C 0 0 0 1 i i i i i i i i i i i i i ii i z d z x a x i i i a a d\u03b8 \u03b1 \u03b8 \u03b8 \u03b1 \u03b8 \u03b1 \u03b8 \u03b8 \u03b8 \u03b1 \u03b8 \u03b1 \u03b8 \u03b1 \u03b1 \u2212 \u2212 \u22c5 \u22c5 \u22c5 \u22c5 \u2212 \u22c5 \u22c5 = =T T T T T (1) Where, S stands for sin , C stands for cos . Simultaneous 0 1T to 7 8T , the homogeneous transformation matrix 0 8T from the base coordinate system 0 0 0x y z to the tools coordinate system 8 8 8x y z is given by equation (2)",
            " In order to analyze the passive joints\u2019 preoperative positioning of the two instrument arms, firstly the instrument and laparoscope arms\u2019 base coordinate systems (BCS) have to be integrated in a general BCS, which is overlapped with the BCS of laparoscope arm; secondly, it is better to place the instrument and laparoscope arms into symmetric position for preoperative positioning operation and arms\u2019 motion, so the two instrument arms are place at two side and the laparoscope arm at the mid place; thirdly, for analyzing the preoperative positioning characteristics, the passive joints of instrument arms are projected on the XY plane along direction of axis Z\u2212 , the equivalent projection links are shown in Fig. 5, the frame G G GX Y Z is the general BCS, frame L L LX Y Z is the laparoscope arm\u2019s BCS, frame A A AX Y Z is the BCS of instrument arm A, and frame B B BX Y Z is the BCS of instrument arm B; point AP is the trocar point projection of instrument arm A, and point BP is the trocar point projection of instrument arm B. According to Fig. 3 and Fig. 5, we can find that it can simplify analysis procedure of passive joints\u2019 preoperative positioning analysis, if the active joints\u2019 values of instrument arms A and B are set as zero, so that the forward kinematics model can be obtained; because of we use symmetric position and in order to achieve maximum space of instrument arms A and B, it means that the passive joints\u2019 values of instrument arms A and B have to be limited within a reasonable range; therefore, these joints\u2019 values or range are set as below: 5 2 2 2 6 3 3 3 7 4 4 4 8 0 [ / 2,0] ' 0 [0, / 2] ' 0 [ / 2, / 2] ' 0d \u03b8 \u03b8 \u03c0 \u03b8 \u03b8 \u03b8 \u03b8 \u03c0 \u03b8 \u03b8 \u03b8 \u03b8 \u03c0 \u03c0 \u03b8 \u03b8 = \u2208 \u2212 = \u2212 = \u2208 = \u2212 = \u2208 \u2212 = \u2212 = (6) According to Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001271_amm.611.279-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001271_amm.611.279-Figure1-1.png",
        "caption": "Fig. 1 The tooth load by norm",
        "texts": [
            " The article is devoted to problems determining of the stress in a dangerous section of tooth foot using FEM. The problem is solved for elliptical, eccentric gear with asymmetrical profile of tooth. The nominally bending stress of gear teeth According to standard STN 01 4686 was calculated bending stress in the foot spur gear teeth for these assumptions. The requirements for accuracy of calculating the resultant force acting on a tooth side effects on the lateral edge a tooth and is introduced into the calculation of impact factor mesh (Fig.1-a), or the resultant force acts on a lonely spot mesh (Fig.1-b). All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-12/07/15,08:47:29) We consider only the bending component load. Dangerous cross-sections for tangential points of the tangents to the transition curves are at an angle of 30 \u00b0 to the axis of the tooth. Calculation of the local bending stress in a dangerous section of the gear tooth where the normal force is applied to the head a tooth (the mesh point A) is computed by equation (1) by [4] \u03b2\u03b5\u03c3 YYY mb F Fa nw t Fn \u22c5\u22c5\u22c5 \u22c5 = (1) In cases where the force acts on a lonely mesh point (point B) bending stress in the dangerous section of the tooth foot calculate by equation (2): \u03b2\u03c3 YY mb F F nw t Fn \u22c5\u22c5 \u22c5 = (2) Where Ft \u2013 the circumferential forse [N], bw \u2013 gear width to calculate the bending [mm], mn \u2013 the module in the normal plane [mm], YFa \u2013 the coefficient of tooth shape [-], YF \u2013 the coefficient of tooth shape [-], Y\u03b5 \u2013 the coefficient of profile mesh impact [-], Y\u03b2 \u2013 the coefficient inclination of the tooth [-]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001604_aieepas.1958.4499867-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001604_aieepas.1958.4499867-Figure4-1.png",
        "caption": "Fig. 4. Singlesuspension bundled assemblies with clamp corona shields",
        "texts": [
            " Shielded double-suspension-string bundled-conductor assemblies. Unshielded Sile-Conductor Assemblies Versus Unshielded Bundled-Conductor Assemblies The evaluation of the performance of bundled-conductor assemblies in this investigation starts by comparing characteristics of a single unshielded-conductor assembly with unshielded bundled assem- blies. The comparisons considered are: 1. A single conductor and bundled con- 90Kaminski, Jr.-Corona Shields for Suspension Assemblies APRIL 195890) O-Unshielded, Fig. 4(A) +-Clamp shield, Fig. 4(B) * -Clamp shield, Fig. 4(C) ~-Clcmp shield, FiS. 4(D) ductors having equivalent current-carrying capacity. 2. A single conductor and bundled conductors, each having approximately one half of the same diameter of the single conductor. The relationship of RIV and corona characteristics for these comparisons is shown by the curves in Fig. 2. A comparison of unshielded bundled-conductor assemblies is shown in Fig. 3. Shielded Single-String BundledConductor Assemblies Bundled-conductor assemblies require the addition of component hardware beyond the requirements of single-conductor assemblies. The influence that such hardware might have on the RIV and corona performance, and the measures which can be applied to improve this performance constitute the major portion of this paper. The technique was to first measure the effect on unshielded bundled-conductor assemblies, and then by isolating the areas of maximum disturbance, apply corrective measures. Fig. 4 illustrates shielding applied to assemblies for limiting the influence of suspension clamps. Fig. 5 shows the influence and effectiveness of shields applied. Fig. 6 includes the addition of a circular 24-inch-diameter shield made from 11/4- inch iron-pipe-size aluminum tubing, used to compensate for the corona and RIV effect of the yoke plate. The re- cADcA, (A, (B) ,.i n0 0 x00 0 4 U, I1-i0 0 Ix (C) (D) Fig. 6. Single-suspension bundled-conductor assemblies with circular shield around yoke plate and clamp shields sults are illustrated by the curves in Fig",
            " Discussion of Data A comparison of the corona and RIV on single and bundled conductors is shown by Fig. 2. The bundled conductors of both diameters (1.05 inch and 1.315 inch) show a substantial decrease in noise level over the 1.9-inch-OD single-conductor assembly. It is of' interest to note that, although the current-carrying capacity of 50 100 150 200 250 300 KILOVOLTS LINE - TO-GROUND Fig. 7. Bundled-conductor assemblies with yoke plate and clamp corona shields, 1.9- inch-OD conductor O-Unshielded, Fig. 4(A) +-Yoke-plate shield, Fig. 6(A) * -Yoke-plate shield and clamp shield, Fig. 6(B) e-Yoke-plate shield and clamp shield, Fig. 6(C) X-Yoke-plate shield and clamp shield, Fig. 6(D) the 1.315-inch-OD bundled-conductor arrangement is 56% greater than the 1.05- inch-OD conductor, no significant difference in RIV level exists between these two assemblies. Since the 1.315-inch-OD conductor represents only a 25% increase in conductor diameter over the 1.05-inch-OD conductor, the ratio of conductor diameters in single versus bundled assemblies will exert greater influence on RIV characteristics than will the ratio of currentcarrying capacities",
            " At voltages higher than 250 kv, the RIV characteristics of the two arrangements are essentially the same. At these higher voltages the increased corona activity on the supporting hardware apparently overcomes the compensating effect of the increased diameter on RIV. The minimum corona points D shown in Figs. 2 and 3 are primarily due to corona activity off the U-bolts of suspension clamps. It follows, therefore, with proper shielding of suspension clamp, improvement in RIV performance can be realized. The three methods employed are shown in Fig. 4. From the curves in Fig. 5 each method employed Fis. 10. Special corona shield to retard undesirable interfering noise shows a definite compensating effect. The saddle shield and circular shield, Figs. 4(C) and (D), are substantially the same with respect to RIV characteristics and are superior to the tube shield of Fig. 4(B). Shields illustrated by Figs. 4(C) and (D) yield a 78% reduction in RIV while the shield in Fig. 4(B) yields a 71% reduction at critical voltage. When corona-onset voltages are taken into consideration the saddle-type shield, Fig. 4(C), shows improvement over the other two types of shields, the minimurm corona-onset point being 277 kv at point B as compared with 245 kv and 250 kv for shields shown in Figs. 4(B) and (D) respectively. Since initial corona onset is observed at the corners of the yoke plate for all assemblies, the saddle shield must afford some grading at this location requiring increased potential stress to initiate breakdown. Through the use of suitable grading shields the corona-onset voltage has been increased but its initial effect has been transferred to the corners of the yoke plate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure4.31-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure4.31-1.png",
        "caption": "Fig. 4.31 Efficiency changes for a given value of the input power. a Projective coordinates, b Cartesian coordinates",
        "texts": [
            " In this case, we have the symmetry of the points C1, B1 relatively to the infinitely remote straight line1 or the base points Q; S. Analogously, the symmetry of the points C2;B2 takes place. 136 4 Two-Port Circuits Similar to (4.74), we may get at once an expression for the effectiveness parameter change m21 AKP \u00bc \u00f0S B1 B2 Q\u00de \u00bc \u00f00 P0\u00f0B1\u00de P0\u00f0B2\u00de 1\u00de \u00bc shc1 shc2 : \u00f04:80\u00de The obtained expression differs from (4.75). From here, it follows that the form of losses change is determined by the initial dependences (4.73) and (4.79). Change of efficiency Let us consider a given value of the power P0 in Fig. 4.31a. This vertical straight line intersects all the hyperbolas with the characteristic effectiveness parameters A = 0, A = 1, A \u00bc 1 and a running value A1. So, we get the corresponding efficiency KP\u00f00\u00de, K1 p , KP\u00f01\u00de, KP\u00f01\u00de in Fig. 4.31b. 4.6 Effectiveness Indices of a Two-Port with Variable Losses 137 Next, we may form the following cross ratio for the value A1, m1 KP \u00bc \u00f0KP\u00f00\u00de K1 P KP\u00f01\u00de KP\u00f01\u00de\u00de: According to (4.79) KP\u00f01\u00de \u00bc \u00f01 P0\u00de; KP\u00f00\u00de \u00bc 1 P0 P0 ; KP\u00f01\u00de \u00bc 1: Analogously, for the value A2, we get the value m2 KP. Similar to (4.77), the efficiency change has the view m21 KP \u00bc m2 KP m1 KP \u00bc \u00f0\u00f01 P0\u00de K2 P K1 P 1\u00de \u00bc K2 P 1 P0 1 K1 P 1 P0 1 \u00bc 1 A2 1 A1 \u00bc sh2c2 sh2c1 : So, there is a strong reason to introduce a specific index in the form KP 1 P0 : Therefore, this index gives more information about a running regime than the simply efficiency"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000043_978-3-030-13273-6_33-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000043_978-3-030-13273-6_33-Figure1-1.png",
        "caption": "Fig. 1. The MRW_4M model.",
        "texts": [
            " 3 computer simulations have been conducted which assume driving through straight line tracks and a track in the shape of a loop, assuming a constant robot frame orientation angle. The article ends with a summary and a bibliography. \u00a9 Springer Nature Switzerland AG 2020 R. Szewczyk et al. (Eds.): AUTOMATION 2019, AISC 920, pp. 346\u2013355, 2020. https://doi.org/10.1007/978-3-030-13273-6_33 In the following section, we analyze the inverse kinematics task of a 4-wheeled mobile robot equipped with mecanum wheels (WMR_4M). When describing WMR_4M kinematics, a model such as that presented in Fig. 1 has been applied. In this representation x, y, z are motionless system axes. The basic components of this model are frame 5 and the driving units. Wheels 1\u20134 are the elements of the driving unit, along with the points located at their centers of symmetry A1;A2;A3;A4, set on semi-axes which are set into motion by the driving module related to a given wheel. These wheels rotate along their own axes which do not change their position relative to the frame. Rollers are located on the wheel perimeter, set at an angle of a \u00bc p=4 [rad] to the driving wheel axis",
            " (18), we will get _xS cos b\u00fe a\u00f0 \u00de\u00bd \u00fe _yS sin b\u00fe a\u00f0 \u00de\u00bd \u00fe _b l cos a\u00fe l1 sin a\u00f0 \u00de \u00bc x2 R\u00fe r\u00f0 \u00de cos a \u00f019\u00de _xS sin b\u00fe _yS cos b\u00fe _bl1 \u00bc xr2r cos a \u00f020\u00de where the coordinates of point S are xS \u00bc xA2 l sin b l1 cos b; yS \u00bc yA2 \u00fe l cos b l1 sin b. Equations (19) and (20) allow us to determine, respectively, the angular rotation velocity of the A2 wheel and that of the roller of this wheel in the requested function of the velocity VS of point S, and the rotation angle of the robot\u2019s frame b t\u00f0 \u00de: Assuming indexes in the markings of the angular velocities of wheels in accordance with their numbering provided in Fig. 1b and proceeding similarly in the case of the other wheels, theWMR_4Mkinematics equations have been determined in the following form: _xS cos b a\u00f0 \u00de\u00bd \u00fe _yS sin b a\u00f0 \u00de\u00bd _b l cos a\u00fe l1 sin a\u00f0 \u00de \u00bc x1 R\u00fe r\u00f0 \u00de cos a _xS cos b\u00fe a\u00f0 \u00de\u00bd \u00fe _yS sin b\u00fe a\u00f0 \u00de\u00bd \u00fe _b l cos a\u00fe l1 sin a\u00f0 \u00de \u00bc x2 R\u00fe r\u00f0 \u00de cos a _xS cos b\u00fe a\u00f0 \u00de\u00bd \u00fe _yS sin b\u00fe a\u00f0 \u00de\u00bd _b l cos a\u00fe l1 sin a\u00f0 \u00de \u00bc x3 R\u00fe r\u00f0 \u00de cos a _xS cos b a\u00f0 \u00de\u00bd \u00fe _yS sin b a\u00f0 \u00de\u00bd \u00fe _b l cos a\u00fe l1 sin a\u00f0 \u00de \u00bc x4 R\u00fe r\u00f0 \u00de cos a \u00f021\u00de Based on the obtained WMR_4M kinematics Eqs. (21) and on the equations of the inverse task of the kinematics, computer simulations have been conducted"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure5-1.png",
        "caption": "Fig. 5 The benchmark of bracket tolerance model a bracket position, b bracket benchmark",
        "texts": [
            " DF-4 is the bottom contact surface between first block and its under bracket; DF-5 is the side contact surface between first block and its upper bracket, which is required to be perpendicular with DF-4; DF-6 (imaginary plane) is the central plane of central hole on first block and brackets, which is required to be perpendicular plane with DF-4 and DF-5 respectively. The under bracket is assembled to rib of VV by contacted both side surfaces of them, which are fixed by four bolts and four location holes. Figure 5a is the under bracket position in VV model, and Fig. 5b is the tolerance model of under bracket including bolts. And its benchmark planes also are defined in here. DF-7 is the upper contact surface of under bracket contacted with third block; DF-8 is the side contact surface of under bracket contacted with rib of VV, which is required to be perpendicular with DF-7; DF-9 (imaginary plane) is the central plane of central hole on under bracket, which is required to be perpendicular plane with DF-4 and DF-5 respectively. Accuracy Requirements of Tolerance Model The ribs and the housing (flexible support for blanket) of VV are welded to inner shell of VV, so their welding tolerance both in dimension and angle are selected with criterion A, B, C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure6-1.png",
        "caption": "Fig. 6 Tolerance requirements of VV ribs",
        "texts": [
            " Accuracy Requirements of Tolerance Model The ribs and the housing (flexible support for blanket) of VV are welded to inner shell of VV, so their welding tolerance both in dimension and angle are selected with criterion A, B, C. Here A, B, C are imaginary criterion plane respectively, and A is toroidal plane, B is poloidal plane, C is radial plane. They are perpendicular reciprocally. So the true positions of DF-1 and DF-2 are 10 by contrast A and B. At the same time, other accuracy requirements are shown in Fig. 6. They are all the accuracy requirements of VV ribs. Except ribs of VV, the accuracy requirements of untagged dimension of blocks and brackets should be executed standard as ISO2768-1, which is \u2018\u2018General tolerance Part 1, Dimension and Angle tolerance for untagged components\u2019\u2019, and class \u2018\u2018f\u2019\u2019 among them are selected as executive standard as Tables 1 and 2. The accuracy requirements of untagged geometric tolerance of blocks and brackets should be executed standard as ISO2768-2, which is \u2018\u2018General tolerance Part 2, Geometric tolerance for untagged components\u2019\u2019, and class \u2018\u2018H\u2019\u2019 among them are selected as executive standard as Tables 3 and 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000526_eeeic.2015.7165465-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000526_eeeic.2015.7165465-Figure2-1.png",
        "caption": "Fig. 2. Geometry of 2-D FEM of the studied IPMSM",
        "texts": [
            " MODELING OF THE DUAL T-TYPE TOPOLOGY CONNECTED TO AN OPEN-END IPMSM It is important to analyze the performance of the advanced topology of dual T-type converters in a complete drive system. Fig. 1-a shows a block diagram for the proposed converter connected to an IPMSM. It consists of two T-type converters, a common dc link for both converters, and a three-phase IPMSM. The converter is supplied by two identical dc power supplies. The circuit diagram for the T-type converter is shown in Fig. 1-b. The geometry of the studied IPMSM is given in Fig. 2. This section discusses the operation of the dual T-type converter. Referred to Fig. 1, the five-level dual T-type converter consists of two identical T-type converters, which will be denoted in the rest of the paper by \u201cconverter-1\u201d and \u201cconverter-2\u201d. It is assumed that all switches, which are a MOSFET type, of converter-2 have the same symbols of converter-1 switches with an additional prime. The dual T-type converter has 8 switches per phase, SiA, and SiA\u2019 with i=1:4. The operation of the dual T-type converter for phase AA\u2019 is summarized in Table I"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.2-1.png",
        "caption": "Fig. 3.2 Different deformation modes of a bending beam: a shear-rigid; b shear-flexible. Adapted from [24]",
        "texts": [
            " \u00d6chsner, Computational Statics and Dynamics, DOI 10.1007/978-981-10-0733-0_3 89 90 3 Euler\u2013Bernoulli Beams and Frames The classic theories of beam bending distinguish between shear-rigid and shearflexible models. The shear rigid-beam, also called the Bernoulli1 beam,2 neglects the shear deformation from the shear forces. This theory implies that a cross-sectional plane which was perpendicular to the beam axis before the deformation remains in the deformed state perpendicular to the beam axis, see Fig. 3.2a. Furthermore, it is assumed that a cross-sectional plane stays plane and unwrapped in the deformed state. These two assumptions are also known as Bernoulli\u2019s hypothesis. Altogether one imagines that cross-sectional planes are rigidly fixed to the center line of the beam3 so that a change of the center line affects the entire deformation. Consequently, it is also assumed that the geometric dimensions4 of the cross-sectional planes do not change. In the case of a shear-flexible beam, also called the Timoshenko5 beam, the shear deformation is considered in addition to the bending deformation and cross-sectional planes are rotated by an angle \u03b3 compared to the perpendicular line, see Fig. 3.2b. For beams for which the length is 10\u201320 times larger than a characteristic dimension of the cross section, the shear fraction is usually disregarded in the first approximation. 1Jakob I. Bernoulli (1655\u20131705), Swiss mathematician and physicist. 2More precisely, this beam is known as the Euler\u2013Bernoulli beam. A historical analysis of the development of the classical beam theory and the contribution of different scientists can be found in [25]. 3More precisely, this is the neutral fibre or the bending line"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000702_1.4031894-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000702_1.4031894-Figure2-1.png",
        "caption": "Fig. 2 Large-displacement buckled shapes for inner-to-outer radii ratios of 0.3, 0.5, 0.7, and 0.84 in (a)\u2013(d), respectively. The mode number, n, is the number of curled \u201cpetals,\u201d equal to 3, 4, 6, and 9. The coloring helps to convey the levels of displacement.",
        "texts": [
            " We consider annuli with b/a in the ratio 0.2\u20130.94, i.e., from wide to very narrow strips, in steps of b=a \u00bc 0:02. Below b=a \u00bc 0:2, there is no difference in behavior; above 0.94, the free edge never properly curls. The outer radius, a, is always 0.1 m, and three thicknesses are specified to give sensible thin shell values for a/t: 0.0001 m, 0.001 m, and 0.005 m. The same two stage asymmetrical analysis described for Fig. 1 is employed again, where the total imperfection amplitude is 0.1%. Some typical postbuckled shapes are given in Fig. 2, where we can clearly see the number of curled petals rising from three for b=a \u00bc 0:3 and beyond. All possess a flattening central region of the same number of sides as curls. The number of curls is taken to be equal to the buckling mode number, which we denote n. For each b/a value, n is plotted in Fig. 3 in terms of the dimensionless width of strip, 1 b=a. Two sets of data are, in fact, plotted, resulting in over 200 data points: for when the inner edge is fully clamped and then repeated for when the in-plane constraint is relaxed",
            " If we assume x and y to be locally aligned to the principal axes of each curl, the twisting curvature, jxy, is zero. In the flat region, jx, jy, and jxy are also zero, but a nonzero jT gives UB \u00bc D\u00f01\u00fe \u00dej2 T. For each curl, UB \u00bc D\u00f01 \u00dej2 T=2. This density is smaller and hence, the annulus will try to maximize the planform area over which curling develops, whilst respecting their compatibility with the central flat region. We may accommodate this requirement by inscribing a polygonal boundary between the regions, as shown in Fig. 5 and evident from the general deformation in Fig. 2. The polygon can be drawn inside the original flat annulus because its interior remains flat for this asymptotic mode of deformation. As n increases, the curling area decreases overall, and thus, fewer curls are favored for a given b/a. The critical value of b/a for a given n occurs when the polygon both inscribes the outer radius, a, and circumscribes the inner radius, b, see Fig. 5(b). If, at this limit, we increase the inner radius slightly, the central hole, which remains flat, now encroaches upon the curls"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000783_s13344-014-0065-9-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000783_s13344-014-0065-9-Figure2-1.png",
        "caption": "Fig. 2. LOS guidance system.",
        "texts": [
            " As the most widely-used, the objective of the LOS guidance system is to keep the vehicle as near as possible to the line of sight between the current positions of the vehicle and the target. For a simple two-dimensional LOS strategy, the surge velocity of the vehicle is assumed to be constant and, considering the line of sight, the desired angle is computed by (Naeem et al., 2003) 1 desired ( ) tan ( ) k k y y t x x t         , (12) where (xk, yk) (k=1, 2, \u2026, N) is the coordinate of waypoint k, and [ ( ), ( )]x t y t is the planar position of the AUV at time t. Fig. 2 shows a schematic diagram of the LOS guidance system. The idea for the development of the 3D guidance system is to adjust the desired depth proportional to the horizontal distance between the AUV and the target. In fact, the guidance system determines the desired depth considering the projection of the distance between the AUV and the target in the horizontal plane to keep the AUV in a trajectory near the straight line connecting the two. With the straight line always close by, the trajectory shall be reasonably short and unsafe areas shall be avoided"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002738_978-3-658-12918-7_25-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002738_978-3-658-12918-7_25-Figure1-1.png",
        "caption": "Figure 1: Test carrier RM2 and floating liner method",
        "texts": [
            " This research engine enables to measure the friction forces of the piston assembly high accurate, crank angle resolved and in a technically relevant field. In a joint research project with Federal-Mogul GmbH specific component tests were performed with the aim to understand and prove the relevant influences of piston assembly design features on the friction force. The base of the research project is the test carrier at the Institute of Internal Combustion Engines, a modular designed single cylinder research engine, shown in Figure 1. The essential design parameters of the engine, like the crank train parameters, correlate to a common passenger car engine. Therefore, it is possible to test components that are close to serial production and to ensure series related testing conditions. The fundamental parameters of the test carrier are summarized in Table 1. The essential element of the test carrier is the measurement device based on the floating liner method, presented schematically in the right side of Figure 1. The cylinder liner is mounted floating on piezoelectric load cells. Thereby, the axial acting friction forces between the piston assembly and the cylinder liner are measured directly and crank angle resolved. The main focuses and challenges in the development of the measurement device were the minimization of interferences, the transferability of the results to series applications and the accuracy of the measurements. The upper location of the load cells and the radial sealing of the combustion chamber are the main technical features"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002436_j.cirp.2015.04.129-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002436_j.cirp.2015.04.129-Figure2-1.png",
        "caption": "Fig. 2. Measuring method to determine the volume VP of preform.",
        "texts": [
            " After total solidification and cooling of the preform, the me VP of the preform is measured as well as the eccentricity e of orm relative to the shaft. For different reasons, the preform does necessarily take ideal spherical shape after solidification. In r to still receive accurate measuring results for the volume of the orm, the length of the rod, which is thermally upset, is taken into unt. This is achieved by marking a horizontal line perpendicular e symmetrical axis of the rod in a specified distance to the lower end (Fig. 2). The distance between the upper end of the preform the horizontal line represents the part of the rod, which is not ten, so that the molten volume can be calculated. An influence of eccentricity on the measuring results is reduced by averaging measurements l1 . . . l4 on the lateral surface of the rod. The surements are carried out using a 3D-microscope Keyence VHX . Assuming that no rod material is vaporized or lost due to ter formation during accumulation process, this method allows easuring accuracy of 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002830_med.2016.7535956-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002830_med.2016.7535956-Figure1-1.png",
        "caption": "Fig. 1. Borea Quadrotor on testbench for single axial testing.",
        "texts": [
            " Based on these preliminary activities, three main contributions are provided in the present paper. The first one is a model identified from experimental data of the involved quadrotor actuators; this model is crucial for control design. The second one is an algorithm for the quadrotor single axial attitude control; the algorithm is based on the Embedded Model Control (EMC) methodology, [3], [4]. The third contribution consists in the presentation of the EMC control results obtained in several experimental tests, carried out on a laboratory single axial testbench, see Fig. 1. To describe these contributions more in detail, consider that a quadrotor has four actuators. Each actuator is composed of three main elements (see also Fig. 2): an electronic speed control (ESC), a motor, a propeller. The motor makes the propeller rotate. This rotation transfers a certain amount of mechanical energy to the air, producing the thrust and torque which are used to drive the quadrotor. The ESC regulates the motor angular rate according to an angular rate reference. Overall, an actuator is a dynamic system with main input the angular rate reference, and main output the propeller thrust"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure1-1.png",
        "caption": "Figure 1. (a) Systems of concurrent springs and dampers under consideration, and (b) equivalent system.",
        "texts": [
            " These approaches yield several theorems and corollaries which enable one to find equivalent systems in a very convenient and straightforward way. In addition, some benefits of using the concept of two mutually orthogonal springs for analysing problems concerned with oscillations of a particle on elastic frames are also demonstrated. Equivalent two-element system for concurrent springs and two-element system for concurrent dampers Let us consider a particle of mass m, which is connected with n arbitrarily positioned linear springs and N arbitrarily positioned linear viscous dampers, as shown in Figure 1(a). Let us also assume that a constant static force F, whose position with respect to the horizontal axis is defined by the angle , acts on the point. As a result of it, the springs are pre-stressed (deformed) in the stable static equilibrium M0 around which the particle performs small in-plane oscillations. Resulting static deformations of the springs are labelled by i (i\u00bc 1,. . ., n), and the stiffness of each of them is ki; the corresponding lengths in the static equilibrium position are li; and the angle between each of them and the horizontal is \u2019i",
            " The dampers are characterized by the damping coefficient cj (j\u00bc 1,. . ., N) and their positions with respect to the horizontal axis are defined by the angle \u2019j. The rectangular coordinates system x\u2013y is introduced in point M0, where the angle defines the position of the x-axis with respect to the horizontal. Of interest here is the replacement of this concurrent system of springs and dampers by the equivalent, but more simple system that consists of two mutually orthogonal springs and two mutually orthogonal dampers, as shown in Figure 1(b). The task is to find their positions, defined by the angles k and c, respectively, as well as the stiffness coefficients kI and kII, and the damping coefficients cI and cII. The way how to find their characteristics is described in the following subsections. Let us consider now the system of concurrent springs described earlier and shown in Figure 1(a) with a view to finding the equivalent system of two mutually orthogonal springs shown in Figure 2(a). The following theorems and corollaries define this equivalent system. Theorem 1. The system of concurrent pre-stressed linear springs performing small in-plane oscillations can be replaced by the equivalent system of two mutually orthogonal springs on which the original constant static force does not act and which do not have the corresponding static deformations. The position of one of them with respect to the horizontal is given by the angle k, which satisfies the following equation tan 2 k \u00bc Pn i\u00bc1 ki 1 i li sin 2\u2019iPn i\u00bc1 ki 1 i li cos 2\u2019i \u00f01\u00de at NORTH CAROLINA STATE UNIV on May 11, 2015pic",
            " The sum of the stiffness coefficients of the springs from the equivalent systems is given by kI \u00fe kII \u00bc Xn i\u00bc1 ki \u00fe ki i li \u00f012\u00de When the springs from the original system are not pre-stressed in the stable static equilibrium position ( i \u00bc 0), the sum of their stiffness coefficients is equal to the one from the equivalent system kI \u00fe kII \u00bc Xn i\u00bc1 ki \u00f013\u00de i.e. the sum of the stiffness coefficients is the invariant of these systems. Proof. It is sufficient to observe that the sum of the equation (2a,b) gives equation (12) and that for i \u00bc 0, this sum yields equation (13). Let us consider now the system of concurrent springs described earlier and shown in Figure 1(a) with the aim of defining the equivalent system of two mutually orthogonal springs shown in Figure 2(b). Theorem 2. The system of concurrent pre-stressed linear springs performing small in-plane oscillations can be replaced by the equivalent system of two mutually orthogonal linear springs on which the original constant static force also acts and which are, consequently, pre-stressed. The position of one of them with respect to the horizontal is given by the angle k satisfying tan 2 k \u00bc Pn i\u00bc1 ki 1 i li sin 2\u2019iPn i\u00bc1 ki 1 i li cos 2\u2019i \u00f014\u00de where ki is the stiffness of each of the original springs, i is their static deformations, li stands for their corresponding lengths in the static equilibrium position and \u2019i is the angle between each of the spring and the horizontal",
            " However, there are other possibilities as well, because the equivalent system contains six unknown quantities k 0 I, k 0 II, lI, lII, I and II that are mutually dependent by means of four equations (17a,b) and (18a,b). Thus, one needs to define two of them a priori and then calculate the rest of them from equations (17a,b) and (18a,b). Corollary 2.1. The sum of the stiffness coefficients of two types of equivalent systems of springs is given by the following relationship kI \u00fe kII \u00bc k 0 I \u00fe k 0 II \u00fe k0I I lI \u00fe k0II II lII \u00f019\u00de Proof. This follows directly from equation (18a,b). Let us consider now the system of concurrent dampers from Figure 1(a) and define the equivalent system of two mutually orthogonal dampers shown in Figure 4. Theorem 3. The system of N concurrent in-plane linear viscous dampers can be replaced by the equivalent system of two mutually orthogonal linear viscous dampers, where the position of one of them with respect to the horizontal is given by the angle c that satisfies tan 2 c \u00bc PN j\u00bc1 cj sin 2 \u2019jPN j\u00bc1 cj cos 2 \u2019j \u00f020\u00de where cj is the damping coefficient of each of the original dampers and \u2019j is the angle between each of them and the horizontal",
            " The system under consideration consists of three linear springs and three linear viscous dampers whose positions as well as stiffness and damping coefficients are shown next to them in Figure 5(a). All the elements are connected in the centre, where the particle of mass m is located. Let us find the equivalent system of two mutually orthogonal springs and two mutually orthogonal dampers. In addition, we will also show how this approach easily gives the natural frequencies of free undamped vibration. Following the notation from Figure 1(a) and using explanations given in Figure 5(b), we identify the angles between the axis of the springs and dampers as \u20191 \u00bc 4 =3, \u20192 \u00bc 2 =3, \u20193 \u00bc 0, \u20191 \u00bc 5 =3, \u20192 \u00bc =3, \u20193 \u00bc . The two-element system of dampers equivalent to the one shown in Figure 5(a) is obtained by calculating first the position of one of the dampers, i.e. the angle c (see equation (20)) tan 2 c \u00bc P3 j\u00bc1 cj sin 2 \u2019jP3 j\u00bc1 cj cos 2 \u2019j \u00bc 2c sin 10 =3\u00f0 \u00de \u00fe c sin 2 =3\u00f0 \u00de \u00fe 3c sin 2 \u00f0 \u00de 2c cos 10 =3\u00f0 \u00de \u00fe c cos 2 =3\u00f0 \u00de \u00fe 3c cos 2 \u00f0 \u00de \u00bc ffiffiffi 3 p 3 \u00f027\u00de the solution of which is c \u00bc 12 \u00bc 15o \u00f028\u00de Equation (21a,b) yields the values of two damping coefficients cI \u00bc X3 j\u00bc1 cj cos 2 \u2019j c \u00bc 2c cos2 5 =3\u00fe =12\u00f0 \u00de \u00fe c cos2 =3\u00fe =12\u00f0 \u00de \u00fe 3c cos2 \u00fe =12\u00f0 \u00de \u00bc 3:866 c, cII \u00bc X3 j\u00bc1 cj sin 2 \u2019j c \u00bc 2c sin2 5 =3\u00fe =12\u00f0 \u00de \u00fe c sin2 =3\u00fe =12\u00f0 \u00de \u00fe 3c sin2 \u00fe =12\u00f0 \u00de \u00bc 2:134 c \u00f029a; b\u00de These two dampers are shown in Figure 6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002572_thc-161219-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002572_thc-161219-Figure1-1.png",
        "caption": "Fig. 1. Equivalent multi-layer geometric model of human forearm.",
        "texts": [
            " The length of human forearm is assumed to be h and equivalent to a multilayer volume conductor cylinder model comprised of skeleton, muscle, fat and skin, (r1, r2, \u00b7 \u00b7 \u00b7 rn) represents circumscribed radius of all tissues on the tangent plane, (\u03b5t1, \u03b5t2, \u00b7 \u00b7 \u00b7 \u03b5tn) and (\u03b5l1, \u03b5l2, \u00b7 \u00b7 \u00b7 \u03b5ln) represent the transverse and the parallel permitivities of all tissues respectively, (\u03b4t1, \u03b4t2, \u00b7 \u00b7 \u00b7 \u03b4tn) and (\u03b4l1, \u03b4l2, \u00b7 \u00b7 \u00b7 \u03b4ln) indicate the transverse and the parallel electric conductivity of all tissues (show in Fig. 1), respectively. In the galvanic coupling human-body communication, when the electrical signal\u2019s frequency in the input electrode is less than 1 MHz [4,7], the propagation effect, the inductive effect and the irradiation effect from the skin to air throughout the channel may be basically ignored. As the frequency is increased, the capacitance effect of the tissue becomes more and more obvious; therefore its impact on the overall system shall be taken into account in building the tissue model. In the cylindrical coordinate system, according to the Maxwell\u2019s equation and the quasi-static approximation conditions [4,5,7], the body surface potential distribution can be approximately shown in the abbreviated Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.5-1.png",
        "caption": "FIGURE 9.5",
        "texts": [
            "2 THIN DOUBLY CURVED ORTHOTROPIC SHALLOW SHELL Nondimensional Center Deflection-Load Curves for CCCC Rectangular Plate E (N/ m )x 2 E (N/m )y 2 G (N/m )xy 2 yx\u00b5 k (1 / m)x k (1 / m)y a (m) b (m) h (mm) 2.0 \u00d7 1010 4.0 \u00d7 1010 1.0 \u00d7 1010 0.1 0.2 0.3 0.1 0.1 0.22 Ex(N/m2) Ey(N/m2) Gxy(N/m2) myx kx(1/m) ky(1/m) 178 CHAPTER 9 GEOMETRIC NONLINEAR ANALYSIS 9.3 BUCKLING OF INCLINED CIRCULAR CYLINDER-IN-CYLINDER 9.3.1 BASIC EQUATION AND SOLUTION PROCEDURES Consider a circular cylinder constrained by an inclined rigid circular cylinder, shown in Fig. 9.5. L, q, and EI are the length, weight per unit length, and bending rigidity of the inner cylinder. The inner cylinder is subjected to a compressive force P at its upper end and a resulting compressive force Fb at its lower end. a is the inclined angle, Wn is the contact force per unit length, and u is the deviation angle Load\u2013Deflection Curves of HHHH Shallow Shell Buckling Mode Shape of HHHH Shallow Shell 1799.3 BUCkLING Of INCLINED CIRCULAR CYLINDER-IN-CYLINDER in the xy plane. The radial clearance between inner and outer cylinders is r",
            ",N ru9kmaxWnk\u22650 u\u00af u\u00af0 u\u00af0 EIuiv+(P+qzs)u0+qzu9+qxu/r=0 EI\u2211j=1N+2Dkju\u00afj+( P+qzsk)\u2211j=1N+2Bkj- u j\u0304+qz\u2211j=1N+2Akju j\u0304 +qxuk/r=0 - (k= 2,3,\u22ef,N\u22121) EI\u2211j=1N+2Ckju\u00afj+(P+ qzsk)\u2211j=1N+2Akju\u00afj= Qk (k=1,N) EI\u2211j=1N+2Dkju\u00afj+(P+qzsk*)\u2211 =1N +2Bkju\u00afj+qz\u2211j=1N+2Akju\u00afj \u22126E I\u2211j=1N+2Akju\u00afj2 \u2211j=1 N+2Bkju\u00afj+qxsinuk/r= 0 (k =2,3,\u22ef,N\u22121) 1839.3 BUCkLING Of INCLINED CIRCULAR CYLINDER-IN-CYLINDER where sk * is measured from the top end of the inner cylinder and varies element by element. Note that =s sk k * is only valid for the first element with node 1 at the top end of the inner cylinder shown in Fig. 9.5. Equation (9.70) should be modified as \u2211 \u2211 \u2211\u03b8 \u03b8 \u03b8      \u2212       + +       = = = + = + = + EI C EI A P q s A Q k N 2 ( ) ( 1, ) kj j j N kj j j N z k kj j j N k 1 2 1 2 3 * 1 2 (9.85) where sk * is measured from the top end of the inner cylinder and varies element by element. For convenient in the assemblage, \u03b8 j should be rearranged as follows, \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8( )= \u2032 = = + \u2032\u2212 +j N, = , 3, 4,..., 1 , =j j N N1 1 2 1 1 2 (9.86) The corresponding equations should be rearranged accordingly, namely, Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000887_gtindia2014-8186-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000887_gtindia2014-8186-Figure1-1.png",
        "caption": "Figure 1. Schematic layout of SFD test rig.",
        "texts": [
            " Analytical studies are carried out to evaluate the critical speed over the tested mass ranges and span length. The parameters of the SFD test rig like oil flow temperature, critical speed, vibration levels and velocities are measured. The test bench is designed to provide a self-contained platform needed for parametric study of squeeze film damping characteristics of the SFD bearing\u2019s used in the Gas turbines. Test setup comprises of the followings. Schematic layout of SFD test rig is shown in Figure 1. Figure 2 shows the installed test rig. Prime mover and step up gear box The prime mover is a squirrel cage AC induction motor, rated 355kW, 3-phase, 2-pole, 50 Hz with TEFC. The motor speed is controlled by a compatible, full-vector, digital drive for operation up to the maximum speed of 2900 rpm. The supports are mounted on an isolated base which is attached to a base frame assembled with motor and a protective guard. A step up gear box (SGB) is connected through gear coupling with motor. This high performance foot mounted (approx ratio 1:7) SGB having the continuous torque of 1140 Nm with the input speed and output speed of 2900 rpm and 20000 rpm respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure3-1.png",
        "caption": "Figure 3 A schematic of torsion test specimen configurations (see online version for colours)",
        "texts": [
            " The machine was controlled using a LabVIEW virtual instrument (VI) where the test rate of angular displacement was maintained at 0.5 rad/min and the corresponding torque could be obtained. Careful calibration was carried out before the tests. The specimens were fabricated in accordance with the ASTM standard E 143 (2002) which is the standard test method for shear modulus at room temperature. The dimension was 25.40 mm (1.00 inch) in diameter with the total length of 152.40 mm (6.00 inch). Specimens with three different configurations were fabricated to obtain the corresponding shear moduli. Figure 3 demonstrates three different specimen configurations TR1, TR2, and TR3. The longitudinal axes of specimen are parallel to x-axis (TR1), at 135\u00b0 to x-axis (TR2), or parallel to z-axis (TR3) of the machine. Each sample had five replicas for torsion tests. The test matrix for torsion test and test results (to be discussed later) are shown in Table 1. An MTS 100KIP tensile test machine was used to perform tensile tests as shown in Figure 4. A 8896.44 N (2000 lbf) external load cell and an MTS 632.85 biaxial extensometer with the accuracy up to 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003238_detc2016-59654-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003238_detc2016-59654-Figure2-1.png",
        "caption": "Figure 2 Contact stresses and crack and subsurface-initiated crack",
        "texts": [
            "51 2 c f f H i f p N              (8) where max and Hp are, respectively, the maximum shear stress due to contact and hydrostatic stress in the subsurface; f  and f  are fatigue strength coefficients for tension/compression and shearing, respectively; and c is a fatigue strength exponent. It is assumed that crack initiation is at a point where the ratio of the maximum shear contact stress to the hardness is maximal and the subsurface crack parallel to the surface is initiated at this point when the number of load cycles reaches iN as illustrated in Fig. 2(a). After the subsurface crack is initiated, the crack propagates under cyclic contact loads to the surface. Using the Paris equation, the crack propagation is modeled by [5]  0( ) ( )m m p da C K K dN     (9) where a is the half length of the crack; N is the number of load cycles; pC and m are constants; and K is the model  stress intensity factor range. 0K is the threshold for the crack growth given by the empirical formula 3 0 2.45 3.41 10K HV    for the Vickers hardness HV [19]. That is, the crack grows only if K is greater than 0K ",
            " It is shown in the literature [7] that growth of the subsurface crack under rolling contact loads is driven by the value defined by a ratio of the maximum shear stress to the hardness, and the following expression for s is suggested by considering the effect of porosity and notch effects [5,7]: 3 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90698/ on 04/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use 2 max( ) Ks HV     (11) where ( 1) 1K tK    and 4.3e   for empirically identified parameters tK ,  and  [21]. Using Eq.9, the number of cycles for the initial crack, which is parallel to the surface as shown in Fig. 2(b), to reach the surface is calculated as  0 0 1 ( ) ( ) ca p m ma p N da C K K     (12) where 0a is the half length of the initial crack as shown in Fig. 2(b) [5]. Accordingly, the total number of load cycles to pitting failure can be predicted by a sum of Eqs. 8 and 12 as a function the maximum shear stress under a cyclic rolling contact load. To account for wind load uncertainty characterized by the averaged joint PDF of the 10-minute mean wind speed (V10) and turbulence intensity (I10) as defined by Eq. 6, multiple 10- minute drivetrain dynamics simulations need to be carried out for various realizations of 10v and 10i , thus the multibody drivetrain dynamics simulation becomes a computational burden in the entire design optimization process"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.15-1.png",
        "caption": "Fig. 4.15 a Holonomic constraint; b nonholonomic constraint",
        "texts": [
            " Holonomic constraints arise from geometric constraints and are integrable into a form involving only coordinates (holo comes from Greek that means whole, integer). Nonholonomic constraints are not integrable. The relation specified by a constraint can be an explicit function of time designated as \u201crheonomic\u201d constraints (rheo comes from Greek that means hard, inflexible, independent) or not, being designated by \u201cscleronomic\u201d constraints (scleros comes from Greek that means flexible, changing). Figure 4.15 shows a typical revolute joint and a simple human body model placed on a spherical surface, which represents a holonomic and a nonholonomic constraint, respectively. Thus, for instance, in the motion of the human model on the spherical surface, the following mathematical relation has to be satisfied during the analysis rTr R2 0 \u00f04:42\u00de where R is the radius of the spherical surface and vector r represents the position of the model measured from the center of the spherical surface. The kinematic constraints considered here are assumed to be holonomic, arising from geometrical constraints on the generalized coordinates"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.32-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.32-1.png",
        "caption": "FIGURE 8.32",
        "texts": [
            " The preload of the spring was 150 lb at the current design, which External force of 150 lb applied to the tire. Spring and damper in the suspension. Physical position of the spring. Shock travel (Case A). FIGURE 8.29 Modified profile cam with a larger hump. FIGURE 8.30 Segment velocity added for Case B. means that the spring was compressed by 1.5 in. when the shock was fully extended. The spring preload was increased from 150 lb (current design) to 200 lb and 250 lb; the resultant shock travels are shown in Figure 8.32. As can be seen, as the preload increased, the overall shock length slightly decreased but the piston moved further away from the cylinder. This is because as long as the compressive force on the shock is larger than the preload, the shock with a relatively large preload will be less compressed than one with a small preload under the action of the same force. The rocker component was parameterized in SolidWorks so that its shape could be adjusted. When the rocker shape changed (Figure 8.33(a)), the overall shock length could be reduced while the piston moved closer to the cylinder (Figure 8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002911_1.4034648-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002911_1.4034648-Figure2-1.png",
        "caption": "Fig. 2 Turbomachinery configuration of interest: (a) threedimensional view and (b) sectional view",
        "texts": [
            " The angular width of the pad is hpad, and the minimum set bore film thickness location within the pad is defined as hoffset. The ratio of these two angles is called the offset ratio c \u00bc hoffset=hpad. The scaling laws are applied to the three-pad foil bearing with various diameters up to 300 mm to investigate whether such a large foil bearing can be designed and adopted to practical industrial applications. The first step is to devise a turbomachinery configuration that may represent typical single stage gas turbine rotors as shown in Fig. 2, where turbine impeller is made of Ni alloy and rest of the system are made of stainless steel. Assuming L/D\u00bc 1, for the nominal bearing size (\u00bc journal diameter) of D, the assumptions in Table 1 were applied to estimate the total rotor weight of the turbomachinery components. The turbomachinery system shown in Fig. 2 was created using the assumptions in Table 1. By suitable scaling of the solid model, the dimensions of all turbomachinery components were obtained for various bearing sizes, and the results of thrust runner outer diameter (OD), impeller diameter, and bearing shaft length are tabulated in Table 2. The mass of individual components and the total load was computed based on the material density of each component. For simplicity, it is assumed that two radial bearings support the half of entire rotor weight equally, and the average bearing pressure was calculated by dividing the half of the total rotor weight by the bearing area (D2)",
            " Typical turbo-air blowers use aluminum impellers instead of steel or Ni alloy impellers, and the motor is located between the two bearings, compensating the weight reduction at the impellers. For applications using process gas lubricants, Organic Rankine Cycle turbo expanders and refrigerant compressors also use aluminum impellers instead of steel impellers, and the motor/generator is located between the two bearings, leading to rather similar rotor weight as the air blowers. Therefore, it is appropriate to use the rotor model in Fig. 2 as a representative system for the applicability study of the scaling laws. The scaling law of support stiffness, Eq. (15), is applicable to any form of support structure. In this paper, corrugated bump foils were chosen as a support structure. For simplicity, it is assumed Journal of Engineering for Gas Turbines and Power APRIL 2017, Vol. 139 / 042502-3 Downloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jetpez/935804/ on 02/20/2017 Terms of Use: http://www",
            " The results from this work serve as a guideline for foil bearing designers in quick choice of initial design parameters. Detailed design of the clearance and support structure stiffness for specific applications requires high level analyses of the bearing characteristics in actual operating conditions, and it is beyond the scope of this study. In addition, rotordynamic stability of the foil bearing-rotor systems depends not only on the bearing itself but also on detailed inertia distribution of the rotor. Investigation of rotordynamic characteristics of the rotor models following the geometry in Table 2 and Fig. 2 is the scope of future study. A0 \u00bc bearing surface area per support structure C \u00bc nominal operating clearance cb \u00bc equivalent viscous damping coefficient D \u00bc diameter of bearing e \u00bc eccentricity E \u00bc Young\u2019s modulus F \u00bc load capacity of bearing fb \u00bc pressure force on structure H \u00bc nondimensional film thickness kb \u00bc stiffness coefficient of support structure L \u00bc length of the bearing N \u00bc shaft speed OBrg \u00bc bearing center P \u00bc nondimensional pressure pa \u00bc atmospheric pressure pavg \u00bc average bearing pressure R \u00bc bearing radius rcg \u00bc centrifugal growth rp \u00bc preload z \u00bc coordinate in axial direction Z \u00bc nondimensional coordinate in axial direction c \u00bc offset ratio \u00f0c \u00bc hoffset=hpad\u00de d \u00bc deflection of the structure e \u00bc nondimensional eccentricity \u00f0e \u00bc e=C\u00de g \u00bc structural loss factor h \u00bc angular coordinate hpad \u00bc angular width of the pad K \u00bc bearing number l \u00bc dynamic viscosity of fluid \u00bc Poisson\u2019s ratio n \u00bc bearing load capacity coefficient q \u00bc density of shaft s \u00bc nondimensional time x \u00bc angular velocity of shaft [1] DellaCorte, C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001494_amm.732.357-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001494_amm.732.357-Figure2-1.png",
        "caption": "Fig. 2 Reissner kinematics of plate",
        "texts": [
            " 00 000 000 \u03b3 \u03b3 \u03b3 \u03b5 \u03b5 \u03c4 \u03c4 \u03c4 \u03c3 \u03c3 \u03c3 L L L L LL LLL E Esym E E EE \u03b5E (1) By the transformation of the system (1, 2, 3) to the system (x, y, z), we obtain the matrix of elasticity in the following form ( ) ( )( ) ( ) ( )\u03b8\u03b8\u03b8\u03b8 TETTETE LL T1T1 == \u2212\u2212 , (2) or ( ) ( )( ) ( ) tt L tt L TETETETE \u03b8\u03b8\u03b8 TT , == , (3) where the transformation matrix has the form ( ) ( ) ( )                 \u2212 \u2212\u2212 \u2212 =      = cs sc scscsc sccs scsc t 000 000 00 002 002 22 22 22 T \u03b8 \u03b8 \u03b8 T0 0T T , (4) 358 Applied Methods of the Analysis of Static and Dynamic Loads of Structures and Machines while ( )\u03b8T satisfies the orthogonality condition ( ) ( )( ) 1T \u2212 = \u03b8\u03b8 TT (5) If we consider the influence of the temperature and the humidity then we get [ ]T21 0,,\u03b1\u03b1=L\u03b1 , [ ]T21 0,,\u03b2\u03b2=L\u03b2 (6) where L \u03b1 is the vector of the thermal expansion, L\u03b2 is the vector of the humidity expansion of the layer with respect to the orthotropic system (1, 2, 3). The transformation of stated vectors is as follows ( ) L\u03b1T\u03b1 \u03b8T= , ( ) L\u03b2T\u03b2 \u03b8T= . (7) The kinematics of the plate with unidirectional reinforced layer according to Reissner-Mindlin theory with the shear consideration is presented in the Fig. 2. The kinematics equations for this theory are as follows. For displacement ( ) ( ) ( )yxzyxuzyxu ,,,, \u03c8\u2212= , ( ) ( ) ( )yxzyxvzyxv ,,,, \u03d5\u2212= , ( ) ( )yxwzyxw ,,, = (8) and for strain x z x u x u x \u2202 \u2202 \u2212 \u2202 \u2202 = \u2202 \u2202 = \u03c8 \u03b5 , y z yy y \u2202 \u2202 \u2212 \u2202 \u2202 = \u2202 \u2202 = \u03d5 \u03b5 vv (9a)       \u2202 \u2202 + \u2202 \u2202 \u2212 \u2202 \u2202 + \u2202 \u2202 = \u2202 \u2202 + \u2202 \u2202 = xy z x v y u x v y u xy \u03d5\u03c8 \u03b3 , \u03c8\u03b3 \u2212 \u2202 \u2202 = \u2202 \u2202 + \u2202 \u2202 = x w z u x w xz , \u03d5\u03b3 \u2212 \u2202 \u2202 = \u2202 \u2202 + \u2202 \u2202 = y w zy w yz v . (9b) The deformation (Eq. 9a, Eq. 9b) can be written in the vector form ( ) ( ) ( )yxzyxzyx ,,,, \u03ba\u03b5\u03b5 += , (10) where Applied Mechanics and Materials Vol"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000612_iecbes.2014.7047531-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000612_iecbes.2014.7047531-Figure4-1.png",
        "caption": "Figure 4. The two link mechanisms represented the kicking leg from overhead view (top), and the example of velocity diagramme drawn (bottom).",
        "texts": [
            " The feet segment was not analysed as the dimension of the foot was small compared to the shank and thigh segment. But, the impact force direction was based on the ankle\u2019s angle at impact, showing that the foot was dorsi-flexed at an angle during the impact. The angle was the flexion angle of the hip while was the knee\u2019s flexion angle. II. The velocity and acceleration diagrammes For the kinematical analysis, the methods of velocity and acceleration diagrammes were utilised also by representing the kicking leg as a two-link mechanism of thigh and shank segment only as shown in Fig. 4. The angular velocity, angular acceleration and absolute velocity and acceleration of the kicking leg were determined using these methods. Table VI shows the velocity table of calculation for the linear velocity and angular velocity required. The velocity of thigh (( was the product of the angular velocity of the thigh and the calculated thigh segment length. The angle and angular velocity of the segment were taken from the measured value from the motion camera. The thigh segment was between the hip (O) and knee (A) joints while the shank segment was between the knee and ankle (B) joints. For this analysis, the hip was assumed as the hinge. The velocity diagramme was constructed based on the two-link diagramme in Fig 4. The link OA underwent circular motion about hinge O at rad/s; the velocity of OA is perpendicular to the link OA. Point A at the link AB also had this velocity, making it the velocity of point B relative to point A. Velocity of OB is the resultant or absolute velocity. The acceleration diagramme on the other hand was constructed based on the similar link segment. However, for the acceleration diagramme, there are radial and tangential components acting towards the origin of the link. The radial component was labelled as an while the tangential component was labelled as at"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002217_icsens.2014.6985116-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002217_icsens.2014.6985116-Figure2-1.png",
        "caption": "Fig. 2. Schematic view of Magnetic Powdery Sensor. The north pole of the magnet is indicated in red and the south pole in blue. Green arrows show the direction of the magnetic field. Grey round objects are iron powder particles. Two magnets are set up in the same direction. The electrodes are SUS304 screws. Sensing area comprises iron powder and magnetic field.",
        "texts": [
            " In this study, we apply this method to sensors and develop a magnetic powdery sensor (MPS). The organization of this paper is as follows. Section II introduces the proposed displacement measurement system. Section III describes the settings and results of a displacement experiment, a repeated displacement experiment, and a repairing experiment. In Section IV, we discuss the results. Section V presents our conclusions and future work. 978-1-4799-0162-3/14/$31.00 \u00a92014 IEEE The MPS is a contact-type displacement sensor that consists of iron powder and magnets. Fig.2 is the schematic view of the MPS. Two sets of ABS plates with a magnet fixed on the center are arranged in parallel. The distance between plates varies from 7 mm to 23 mm, and their magnetic fields are in the same direction. We implement the MPS by dispersing iron powder between the plates. Iron powder is attracted to a magnetic field and lines up along the field (Fig.3). When the distance between the two plates is changed, the diameter of the sensing area changes (Fig.4). This diameter change causes a change in the value of the resistance of the sensing area"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000675_1.4919120-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000675_1.4919120-Figure2-1.png",
        "caption": "FIG. 2. Build up of magnetic indicator. (a) shows the initial / relaxed state with liquid solution and (b) the state after activation with solid solution.",
        "texts": [
            " The proposed sensor consists of two main components, (i) an RFID tag for easy access as well as good usability and (ii) a magnetic temperature threshold indicator, as shown in Fig. 1. With the RFID tag, it is possible to achieve sensor identification (ID) using ISO15963 for communication which is also used in common smart phones with near field communication technology. Hence, it is possible to read such sensors with a smart-phone which allows easy verification of marked goods. To acquire indicator functionality, a non-reversible measurable change due to a threshold temperature has to be achieved. The threshold indicator consists of three magnets, as shown in Fig. 2. One torus shaped magnet in advance referred as indicator and two cuboid shaped magnets. The magnets are sintered hard ferrites. The cuboid magnets are uniaxial magnets, with an anisotropy axis in x-direction. Hence, the only stable magnetization state is in the 6x-direction. The torus shaped magnet is an isotropic magnet, which can be magnetized in any direction. For material information, see Table I. All magnets are mechanically fixed, but the indicator is allowed to rotate freely. For this prototype, a polyamid 6 casing was used",
            " Sensor activation is possible if the ambient temperature is long enough below the threshold temperature for the solution to become solid. (iii) If the solution is solid (frozen) it is possible to magnetize the indicator perpendicular to the cuboid magnets with a magnetic field pulse. Due to the higher coercive field and the fact that the cuboid magnets are uni-axial magnets with easy axis in x-direction, they keep their magnetization in x-direction. However, the indicator (torus shaped magnet) becomes magnetized close to the y-direction, as shown in Fig. 2. The field pulse for activation needs to be very specific in direc- tion and strength, which results in a high security againsta)Electronic mail: roman.windl@tuwien.ac.at 0021-8979/2015/117(17)/17C125/4/$30.00 VC 2015 AIP Publishing LLC117, 17C125-1 manipulation. Now the sensor is activated and ready for monitoring purposes. Since the magnetization direction of the indicator is not parallel to the magnetization of the cuboids, a torque is exerted on the indicator. The indicator rotates towards the magnetization direction of the cuboids only if the threshold temperature is exceeded long enough for the solution (or metal alloy) to melt"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003009_ems.2015.38-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003009_ems.2015.38-Figure1-1.png",
        "caption": "Fig. 1 Scheme of four mass system",
        "texts": [
            " This part may be depicted, in a general case, as an multi-mass system [10]. In this paper four \u2013mass system was considered. Transition delay of the torque control loop was assume 0.3 mssam\u03c4 = and inertia of the torque control loop 0.2 mscur\u03c4 = . Torque constant is equal to 17.5 Nm/ATk = . PMSM (permanent-magnet synchronous motor) was considered. In this article has been considered a four-mass system with one motor generating torque MT and one load torque LT on opposite side. Schematic diagram of the four mass mechanical system was shown in Fig. 1. In the simulation tests transfer functions models were used, which were connected as shown Fig. 2. The derivation of multi-mass mechanical model was described in paper [10]. Model of drive with flexible multi-mass connection can be represented by transfer function (2), where J\u03a3 is total moment of inertia and 3RL = is a number of resonance blocks described by (3). 978-1-5090-0206-1/15 $31.00 \u00a9 2015 IEEE DOI 10.1109/EMS.2015.38 195 ,1,1 1 ( ) 1 1( ) ( ) RL r i i M M ss G s J s H T \u03c9 \u03a3 = = \u22c5 \u22c5= \u220f (2) The angular frequencies 2r rf\u03c9 \u03c0= , 2ar arf\u03c9 \u03c0= define location of resonance and antiresonance in the Bode characteristics"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000681_s40313-015-0208-0-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000681_s40313-015-0208-0-Figure1-1.png",
        "caption": "Fig. 1 Illustration of the path-following problem",
        "texts": [
            " All these features constitute the contribution of this paper relative to recent others in the same scope. The paper is organized as follows. Section 2 states the problem that will be addressed, and Sect. 3 presents the proposed control scheme. In Sect. 4, real results are presented showing the performance of the original and the proposed control scheme. Finally, the conclusions are drawn in Sect. 5. The path-following problem is normally characterized through parametrizations in terms of a path length s, unlike the approach called trajectory tracking, in which the path is parameterized by time. Figure 1 shows the nonholonomic vehicle, the coordinates systems, and a reference path \u0393 . The kinematic model of mobile robots can be defined relative to a Serret\u2013Frenet system {F}, moving along a reference path. This formulation uses the Serret\u2013Frenet system to model the movement of a virtual vehicle that should be followed by the real vehicle. The pose of the virtual vehicle is defined by a vector of reference states [ xr yr \u03b8r ]T . The error vector between the states of the real and virtual robots relative to {F} is given as follows: x{F} e = R(\u03b8r )x{W } e ; x{F} e = \u23a1 \u23a3 xe ye \u03b8e \u23a4 \u23a6 = \u23a1 \u23a3 cos \u03b8r sin \u03b8r 0 \u2212 sin \u03b8r cos \u03b8r 0 0 0 1 \u23a4 \u23a6 \u23a1 \u23a3 x \u2212 xr y \u2212 yr \u03b8 \u2212 \u03b8r \u23a4 \u23a6 . (1) A detailed analysis of Fig. 1 shows that using the error vector (1), a forward velocity v and the kinematic model of the robot (Soetanto et al. 2003), the path-following control problem can be defined as a regulatory control problem. The model of the state error relative to the coordinate system of the path is defined as: x\u0307e = yek(s)s\u0307 \u2212 s\u0307 + v cos \u03b8e; y\u0307e = \u2212xek(s)s\u0307 + v sin \u03b8e; \u03b8\u0307e = \u03c9 \u2212 k(s)s\u0307, (2) where k(s) is the curvature in the section limited by length s. Therefore, the path-following problem is solved through the attainment of feasible values to s\u0307 and \u03c9"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002029_iemdc.2015.7409101-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002029_iemdc.2015.7409101-Figure10-1.png",
        "caption": "Fig. 10. Calculated (FEM) iron loss density per unit value distribution in the stator core of tested motor in Operation mode I [kW/m3].",
        "texts": [
            " At 4000 min-1 10 Nm, efficiency in Operation mode II is about 2 % increase rather than that in Operation mode I (Fig. 7 (b)). Efficiency improvement effect comes large when torque goes lower. Fig. 8 and Fig. 9 show calculated motor loss versus torque of tested motor by FEM in Operation mode I For the same torque, Operation mode II needs about twice phase current rather than Operation mode I. As discussed in section II, in low torque region, PWM carrier harmonic iron loss is not increase if phase current become double. Fig. 10 and Fig. 11 show calculated (FEM) iron loss density per unit value distribution in the stator core of tested motor for Operation mode I and Operation mode II, respectively. Calculated condition is at speed 4000 min-1, torque 10 Nm, DC voltage 400 V, and PWM carrier frequency 15 kHz. In Operation mode I (Fig. 10), iron loss density distribution is equally exist every teeth and back TABLE II CALCULATED CONSUMED POWER AND IMPROVEMENT RATE IN JC08 MODE DRIVING CYCLE. yoke. On the other hand iron loss density distribution on Non-excitation teeth and half of back yoke is reduced in Operation mode II (Fig. 11), though, that on excited teeth is almost equal to that in Operation mode I. From these results, it is confirmed that carrier harmonic current ic isn\u2019t increase even if phase current increase as long as the hypothesis in Consumed power energy (Hybrid mode) 451"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002099_red-uas.2015.7441020-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002099_red-uas.2015.7441020-Figure14-1.png",
        "caption": "Fig. 14: Avionics.",
        "texts": [],
        "surrounding_texts": [
            "In this paper, we have presented a convertible unmanned aerial vehicle (CUAV) whose main characteristics is to perform hover and cruise flight. The dynamic model was obtained using the Newton-Euler approach and it considers the aerodynamic effects. For control of the complete operation of the CUAV, we have proposed an adaptive backstepping control, based on quaternions, which guarantees stability under a singularity-free representation. Numerical simulations showed the performance of the aerial vehicle and the stabilization of the closed-loop system. The CFD simulation allowed us to obtain a performance of the vehicle in cruise mode. Finally, we have described the low-cost avionics of the CUAV for the complete operation."
        ]
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure8.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure8.10-1.png",
        "caption": "Fig. 8.10 Mapping of reference triangles",
        "texts": [
            "40), we find the subsequent currents I21 \u00bc qI21 q1 \u00bc m21 1 I11 m21 1 1 IG11 I11 \u00fe m21 2 1 IG22 I12 \u00fe 1 ; I22 \u00bc qI22 q1 \u00bc m21 2 I12 m21 1 1 IG11 I11 \u00fe m21 2 1 IG22 I12 \u00fe 1 : \u00f08:42\u00de I23 \u00bc qI23 q1 \u00bc J 21 11I 1 3 \u00fe J 21 12I 1 4 \u00fe J 21 13 J 21 31I 1 3 \u00fe J 21 32I 1 4 \u00fe J 21 33 ; I24 \u00bc qI24 q1 \u00bc J 21 21I 1 3 \u00fe J 21 22I 1 4 \u00fe J 21 23 J 21 31I 1 3 \u00fe J 21 32I 1 4 \u00fe J 21 33 : \u00f08:43\u00de We note that the denominators of expressions (8.42) and (8.43) are equal to each other. 8.2.4 Two Cascaded Four-Port Networks Let us consider cascaded four-port networks in Fig. 8.9. The first four-port is given by a matrix Y3 6 of Y parameters and the second four-port corresponds to a matrix Y1 4. 252 8 Passive Multi-port Circuits Similarly, we may superpose the system of coordinates \u00f0I5 0 I6\u00de with the systems of coordinates \u00f0I3 0 I4\u00de, \u00f0I1 0 I2\u00de in Fig. 8.10. Then, a projective transformation, which transfers points of the plane \u00f0I1; I2\u00de into points of the plane \u00f0I5; I6\u00de, takes place. Therefore, the reference triangle G10G2 corresponds to the triangle ~G1~0~G2. Also, a unit point SC, running regime point M1 correspond to points ~S~C; ~M1. Moreover, two bunches of the straight lines \u00f0I1; I2; YL1\u00de \u00bc 0; \u00f0I1; I2;YL2\u00de \u00bc 0 correspond to two bunches of the lines \u00f0I5; I6; YL1\u00de \u00bc 0; \u00f0I5; I6;YL2\u00de \u00bc 0 with centers in the points ~G2; ~G1. Also, the point ~M1 is defined by projective non-uniform and homogeneous coordinates which are set by the reference triangle ~G1~0~G2 and a unit point ~S~C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000511_med.2015.7158890-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000511_med.2015.7158890-Figure12-1.png",
        "caption": "Fig. 12. Paths computed for grasping the object. In red circles, the path calculated with DE. In blue, the path computed with gradient descent. The green lines are the representation of the palm and the finger\u2019s phalanges. The black dots correspond to the joints and the fingertips. The cube in gray, is the object to be grasped. The yellow lines represent a triangle formed by the grasping positions of the fingers.",
        "texts": [
            " X = [\u03b1,\u03b2 ,\u03b3] i = [x(1),x(2)] = FK f inger(\u03b1,\u03b2 ,\u03b3) cost = { max(D) if dE(i,bestk\u22121)= 1 D(i) otherwise (8) where X is the parameter vector of the optimisation process, f k f inger is the forward kinematics function of the finger, bestk\u22121 is the solution of the last iteration of the DE algorithm, dE is the euclidean distance function and D is the funnel potential of the fingertip. It is important to note that the path for each finger is computed separately, so the algorithm is executed once per finger. Figure 12 shows the paths computed for the fingers in order to achieve the grasping positions. It can be appreciated that the paths calculated with FM2 and DE are much more ergonomic. The algorithm presented introduces a new manner of dealing with restrictions in the Fast Marching Square (FM2) method for path planning. This approach has been tested introducing soft and hard restrictions in a single bar, a double bar and a hand-like systems. We have obtained satisfactory results for the three of them using two different search spaces in the path extraction: cartesian and angular varaibles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003276_physreve.94.063002-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003276_physreve.94.063002-Figure1-1.png",
        "caption": "FIG. 1. Elastic inversion of a thin conical shell using finite element analysis. The initial cone angle, measured from the vertical, is 60\u25e6 and the side length of cone is 50 mm; the apex is a small spherical cap of radius 3 mm and everywhere the thickness is 0.1 mm. (a) Initial configuration where a force is applied to the apex and the base of the cone is held rigid. (b) Two views of the partially inverted state, highlighting contours of maximum principal strain. These are focused in the original apex and around the ridge that separates the inverted part from the original upright part. (c) Cutaway view of (b) from afar and close up, showing the deviation of the conical shape near the ridge line away from perfectly straight meridians (gray lines). The width of this deviation is roughly the span over which the strain contours diminish in variation and is shown by the arrow: This is the conical crease associated with the ridge line. (d) Almost fully inverted cone, indicating the extent of deformation during analysis.",
        "texts": [
            " They revise upward the number of latitudinal regimes to seven, with five alone describing the inverted shape between the central loading point and the ridge, followed by the ridge itself and the original part beneath. These regimes are found from an elegant asymptotic analysis *kas14@cam.ac.uk of the governing equation of axisymmetrical deformation, which is compared against its full numerical solution and finite element analysis. There is a richness of results; however, the crease extent is expressed only in order-of-magnitude terms because it is not of explicit interest. Here we examine the precise shape of a similar crease formed during inversion of a right circular cone. Figure 1 indicates some partially inverted shapes obtained from the finite element analysis of the next section. Three major geometrical regimes are evident: upright and inverted sections, which are largely undeformed, and a narrow crease. We can surmise a Pogorelov-type profile in the limit of zero thickness, where the cutaway in Fig. 1(c) shows the more heavily strained area pertaining to the crease diverging from this profile superimposed. Axisymmetry is prescribed at all times during computation even though new studies suggest the possibility of secondary buckling of the circular ridge into a rough polygonal outline in practice [4]. We will be concerned only with solving for the shape of an axisymmetrical crease using the linear governing equation of deformation for a cylindrical shell. Such an approach would appear to flout the major effects of geometrical nonlinearity and initial cone shape upon the crease shape, but it greatly reduces the complexity of analysis and enables accurate closed-form solutions provided two key assumptions are upheld",
            " Under a central point force, this region locally inverts early on in the deformation profile, enabling conical inversion to become well developed, the starting point of our study. We consider a range of cone angles, including a tube in the extreme, and we highlight representative behavior, in particular, our definition of crease width according to how the ridge strains attenuate. We then present our analytical model and compare predictions with finite element data before concluding with a brief discussion. The commercial software package used in Fig. 1 and throughout is ABAQUS [7]. Elements are axisymmetrical twonode linear SAX1 elements available from the standard library and each conical model has 500 elements along the defining meridian; trials using fewer three-node quadratic elements makes little difference to the computational efficiency. The material has a typical Poisson ratio of 0.3 and a Young modulus of 1 MPa, which is relatively soft but intends to mimic a soft rubber, and all thicknesses are of the order of 0.1 mm for a cone side length of around 50 mm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003932_pi-4.1952.0037-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003932_pi-4.1952.0037-Figure3-1.png",
        "caption": "Fig. 3.\u2014Graphical determination of critical angle.",
        "texts": [
            " At some angle 92, greater than 9X, (d92/dt)2 must vanish, which is to say that ST must touch or intersect curve (b); furthermore, we must have l12 But PM/co is the slope of ST and (r^E^coX^) sin 92 is the slope of curve (b) at 92, so the tangent condition will fulfil both requirements. The graphical method of determining the critical clearing ALTERNATORS: A GRAPHICAL SOLUTION OF THE TWO-MACHINE CASE 369 (\u00ab) -]~~ (cos Oo - cos 6). (/>) 2 \u00a7 r ^ (cos Oo - cos 6). angle may be described with reference to Fig. 3 as follows. Plot the two cosine expressions of eqn. (10) as curves (a) and (b). From the point P on curve (a) draw the straight line PQ of slope PMla) and draw also ST parallel to PQ and tangential to curve (b). With the upper curve as base, set up some values of the lower intercept y and join them on a curve. The intersection of this curve with ST defines the critical clearing angle 0,. (2.1) Determination of the Time Scale The graphical method provides a simple means of relating time and angular displacement, for intercepts such as y give the value of (I/2)(dd/dt)2 and the abscissa of the graph is 6",
            " The total time from 6 =\u2022= 0 to 6 = 6r is therefore The critical clearing time may therefore be determined. In previous treatments of the problem4 the clearing time has been derived from a double integration of eqn. (11); in the method described here the time may be obtained by a single step-by-step operation on eqn. (12)\u2014or on the graph representing it\u2014under favourable conditions as regards accuracy and ease of calculation. It is seen from eqn. (13) that the square root of the intercept is used, thus reducing by some 50% the effect of errors in the measurement of y, whilst from Fig. 3 it is clear that considerable lengths of the curves (a) and (b) may be treated as straight lines, thus reducing the number of intercepts VOL. 99, PART IV. that must be measured. Furthermore, errors made in measurement and calculation are not cumulative in the final result, as they are for integration methods. The graphical method has been tested by a calculation of the critical switching time under various conditions of initial load transfer for a system in which r{ and r2 of eqn. (10) are respec- \u2014 x \u2014 x \u2014 Graphical results"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002269_iccas.2014.6987753-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002269_iccas.2014.6987753-Figure1-1.png",
        "caption": "Fig. 1 Preassigned regions of desirable roots location.",
        "texts": [
            " To pose such a problem mathematically, let introduce certain functional J = J(W) = [Yet, W), 8(t, W), u(t, W)] , (5) which is given on the motions of the closed-loop nonlinear system, presented by Eqs. (1), (2), (4). Let us also introduce an admissible set of the controllers (4) to be considered bellow as follows: Qw={ W: OJ(W)EC,,, i=l,nd } . (6) Here 0; is the i - th root of the characteristic polynomial Ll(s) of the closed-loop linear system (3), (4); nd = degLl ; C\", is a preassigned region of desirable roots location on the complex left-half plane. Fig. 1 presents examples of the most popular variants of the regions C\", , where required stability degree a > 0 is taken into account. We pose the problem of analytical synthesis a stabilizing control law (4) for the plant (3) as the following optimization problem J = J (W) \ufffd min , (7) WEQW supposing that the greatest lower bound of the functional on the set Qw can be achieved here. Remark that the problem posed above seems to be extraordinarily complicated one; this motivate us to find its simplified representations, which are convenient for practical using",
            " Otherwise, we shall say that the structure is incomplete. For any functional J = J(W(s,h\u00bb) = / (h) given on the motions of the mentioned system we have the following conversion of the problem (7) /(h) \ufffd min , (10) hEQh where Qh = {h E EP : W(S,h)E Qw} . Here the set Qw has the same sense as (6). Defini tion 2: The problem (10) is called a problem 0/ modal parametric optimization. Let characteristic polynomial of the closed-loop connection be A(s,h) , nd = degA(s) . We accept now that a desirable region for its roots assignment is C \ufffd = C \ufffd3 (Fig. 1 c) as the most general case for the variants presented above. Suppose that the curve bounding this region is determined by the real continuous function 'V='V(v) \ufffd 0 , VE [0,00) , such that '1'(0) =0 . To assign desirable poles of the system, we must satisfy the identity A=: A* , where polynomial A* can be constructed on the base of the following statement: Theorem 1: For any vector yE En d all the roots of the auxiliary polynomial * { \ufffd* (s, y), if nd is even; A (s, y) = \ufffd* (s + an 'l+l (Y,a\u00bbA (s, y), if nd is odd, (11) belong to the region C \ufffd = C \ufffd3 (Fig. 1 c), and conversely, if all the roots of some polynomial A(s) belong to the mentioned region, then there exists the vector yE En \" such that the identity A(s):= A* (s, y) holds. Here we use the following notations: nq = [nd 12], ,--...,* n q 2 1 0 A (s,Y) = IT(s +ai(y,a)s+ai (y,a\u00bb) , (12) i=! /(-):(-00, +00)\ufffd (0,1) IS any function, having reciprocal one for every point of its definition. Proof: A validation of the claim directly follows from the elementary features of quadratic trinomials in the formulae (11), (12)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000702_1.4031894-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000702_1.4031894-Figure5-1.png",
        "caption": "Fig. 5 Inextensible model for calculating mode number, n. (a) Initial planform geometry where b and a are inner and outer radii, respectively. The inscribed polygon (shown here as a triangle) has n sides. Inextensibility implies that the displacement field can only be cylindrical or flat: inside the polygon, the disk remains flat, and outside (in a typical hatched region), the disk is cylindrically deformed. (b) The inscribed polygon has side-length 2l1, and just touches the inner radius; when the inner radius is slightly larger (dashed) then the order of polygon must increase to avoid encroachment, to satisfy displacement compatibility. (c) The limiting inner radius when the polygon, of side-length 2l2, is just touching.",
        "texts": [
            " If we assume x and y to be locally aligned to the principal axes of each curl, the twisting curvature, jxy, is zero. In the flat region, jx, jy, and jxy are also zero, but a nonzero jT gives UB \u00bc D\u00f01\u00fe \u00dej2 T. For each curl, UB \u00bc D\u00f01 \u00dej2 T=2. This density is smaller and hence, the annulus will try to maximize the planform area over which curling develops, whilst respecting their compatibility with the central flat region. We may accommodate this requirement by inscribing a polygonal boundary between the regions, as shown in Fig. 5 and evident from the general deformation in Fig. 2. The polygon can be drawn inside the original flat annulus because its interior remains flat for this asymptotic mode of deformation. As n increases, the curling area decreases overall, and thus, fewer curls are favored for a given b/a. The critical value of b/a for a given n occurs when the polygon both inscribes the outer radius, a, and circumscribes the inner radius, b, see Fig. 5(b). If, at this limit, we increase the inner radius slightly, the central hole, which remains flat, now encroaches upon the curls. This is incompatible for the present polygonal layout, and so n must increase. This cross-over value is found by identifying two side-lengths of polygon: 2l1 in Fig. 5(b) for the inscribing polygon and 2l2 in Fig. 5(c) for the circumscribing polygon. The interior angle, p=n, is also identified in Fig. 5(a). Simple geometry tells us: sin p=n \u00bc l1=a and tan p=n \u00bc l2=b, and the critical condition therefore has 2l2 \u00bc 2l1, that is cos p=n \u00bc b=a for n 3 (2) This locus is plotted in each of the previous subfigures in Fig. 3. At low n, it follows the same trend but diverges from all data beyond n\u00bc 6. When plotted on logarithmic axes in Fig. 4, it has different slope from the finite element data, equal to about 2. This is evident when we consider p=n to be small in Eq. (2) and then expand using the Binomial Theorem 1 b=a \u00bc 1 cos p=n p2 2n2 (3) This \u201cpacking\u201d model works reasonably well by showing that n must increase as the annular width falls, but it does not capture the general trend accurately enough"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003238_detc2016-59654-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003238_detc2016-59654-Figure3-1.png",
        "caption": "Figure 3 Gear coordinate systems",
        "texts": [
            " For this reason, amount of tip relief is one of the important design parameters together with the tooth face width, and the effect of micro geometry needs to be precisely evaluated using the gear dynamics simulation. For this reason, a numerical procedure for multibody gear dynamics simulation based on the tabular contact search algorithm [24] is introduced and integrated into the gear design optimization procedure considering wind load uncertainty in this study. This procedure allows for detecting the gear tooth contact in an efficient manner by introducing the look-up contact tables while retaining the precise contact geometry and mesh stiffness variation in the evaluation of mesh forces. As shown in Fig. 3(a), the global position vector of a contact point on the tooth k of rigid gear body i is defined by [24] ik i i ik r R A u (13) where iR is the global position vector of the origin of the body coordinate system; iA is the rotation matrix; and iku defines the location of the contact point defined with respect to the body coordinate system and is defined as 1 2 0 0 1 2( , ) ( , )ik ik ik ik ik ik ik ik ps s s s u u A u (14) where 0 iku and 0 ikA define the location and orientation of the tooth profile coordinate system with respect to the body coordinate system, respectively. Since the gear tooth contact is periodic, solution to the contact geometry problem of one-tooth contact model as shown in Fig. 3(b) can be repeatedly used for 4 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90698/ on 04/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use detecting the contact of all the teeth for a single pair of gears. For this reason, the look-up contact tables obtained for the onetooth model are generated using the contact geometry analysis and then they are used to determine the location of the contact point in the dynamic simulation. To find the location of contact point k for gear surface i and j, the following non-conformal contact condition is imposed as shown in Fig. 3(c) [22,24,26- 28]: 1 2 1 2 ( ) ( ) ( , , , ) ( ) jk ik jk jk ik jk ijk i j ik jk jk ik jk ik jk ik jk                   t r r t r r C q q s s 0n r r t n t n (15) where 1 1/ik ik ik p s  t u , 2 2/ik ik ik p s  t u , and 1 2 ik ik ik n t t at contact point k on surface i defined with respect to the body coordinate system. In the case of measured tooth profiles with tooth surface imperfections, finding the correct contact point becomes difficult due to the occurrence of jumps in contact on an irregular surface [24]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.19-1.png",
        "caption": "Fig. 35.19 Spherical roller thrust bearing.",
        "texts": [
            " Two and four row taper roller bearings are also made for applications such as rolling mills. Thrust ball bearings are designed to accommodate axial loads. They are not suitable for radial loads. To prevent sliding at the ball to raceway contacts, caused by centrifugal forces and gyratory moments, thrust ball bearings must be subjected to a certain minimum axial load. The bearings are of separable design and the housing and shaft washers may be mounted independently. Spherical roller thrust bearings (Fig. 35.19) In spherical roller thrust bearings the line of action of the load at the contacts between the raceways and the rollers forms an angle with the bearing axis, and this makes them suitable for carrying a radial load. This radial load must not exceed 55% of the simultaneous acting axial load. The sphered raceway of the housing washer provides a selfaligning feature which permits, within certain limits, angular displacement of the shaft relative to the housing. A rotating machine element, e.g. the shaft, generally requires two bearings to support and locate it radially and axially relative to the stationary part of the machine, e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003650_9781782421955.699-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003650_9781782421955.699-Figure2-1.png",
        "caption": "Figure 2 \u2013 FZG test machine.",
        "texts": [
            "4 - - Silicium (Si) [ppm] - - 3 19 Molybdenum (Mo) [ppm] - - 12 1150 Sulphur (S) [ppm] 11200 5020 6265 1800 Physical properties Density @ 15 \u00b0C [g/cm3] 0.902 0.859 0.863 0.861 0.855 Thermal expansion coefficient(\u03b1t\u00d710-4) [/] -5.8 -5.5 -7.0 -7.4 -7.5 Viscosity @ 40 \u00b0C [cSt] 319.2 313.5 332.65 310.07 307.75 Viscosity @ 80 \u00b0C [cSt] 43.9 60.4 91.17 86.31 92.41 Viscosity @ 100 \u00b0C [cSt] 22.3 33.3 39.25 31.98 30.50 m [/] 9.066 7.351 7.134 7.302 7.238 n [/] 3.473 2.787 2.698 2.767 2.739 VI [/] 85 150 159 152 150 3 TEST RIG Figure 2 presents the FZG test machine used for this work. This gear test rig uses the re-circulating power principle [9]. The test pinion (1) and wheel (2) are connected to the drive gearbox by two shafts (3). The shaft connected to the test pinion (1) is divided into two parts by the load clutch (4). One half of the clutch can be fixed with the locking pin (5), whereas the other can be twisted using the load lever and different weights (6). The torque loss ( ) was measured using an ETH Messtechnik DRDL II torque transducer assembled on FZG test machine"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000996_ijsurfse.2014.060481-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000996_ijsurfse.2014.060481-Figure2-1.png",
        "caption": "Figure 2 Schematic diagram of the laser beam travel",
        "texts": [
            " The commonly used approach to model the heat flow induced by a distributed heat source moving over a semi-infinite planar solid starts with the solution for a point source and integrating it over the area of the beam (Cline and Anthony, 1977). Since this method requires complex computations, the analytical model developed by Ashby and Easterling (Ashby and Easterling, 1984; Ahmed et al., 2010; Selvan et al., 1999; Chiang and Chen, 2005), was used in this study to obtain an approximate solution. The coated strip is a fixed solid workpiece over which a laser beam of power q and radius rx moves with velocity v in the x-direction as shown in Figure 2. A point at (y, z) below the surface is thus subjected to a thermal cycle T(y, z, t). If we assume that the solid is large enough such that the average temperature T0 is unchanged, then the absorbed laser energy heats the surface, which is quenched by conduction. According to the Shercliff-Ashby-Easterling (SAE) model, the temperature field around a moving Gaussian shaped beam is given by ( )( ) ( )2 20 0 1/2 00 1( , , ) exp 4 z zAq yT y z t T t t t\u03c0kv t t t \u239b \u239e+ \u239c \u239f\u2212 = \u2212 + \u239c \u239f++ \u239d \u23a0\u03b1 (1) The constant t0 is the time it takes for heat to diffuse laterally over a distance equal to half the beam width in the other direction, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000859_1.4027130-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000859_1.4027130-Figure4-1.png",
        "caption": "Fig. 4 Initial positions of the geometric and rotation centers of a single-row trochoidal gear (Type A)",
        "texts": [
            " In the experiments, the roller gear and cam gear were mounted on the input and output shafts. For preloading the single-row trochoidal gears (types A and B gears), the center distance Cd between the input and output shaft was set at 192.99 mm, which was 10 lm less than the theoretical center distance (193 mm) for nonpreloading. Since the initial locations (locations at the start of operation) of Oin, Oout, Or, and Oc may affect the transmission errors in the experiments, the initial locations were set as shown in Fig. 4, in which Oout-XY is the coordinate system at rest and b is the angle between the X-axis and xc-axis. In the initial locations, Or, Oin, and Oout were in-line. The values b for types A and B gears are given in Table 2. Table 1 Basic specifications of single-row trochoidal gears (Types A and B) Type A Type B Number of rows 1 1 Number of rollers zr 10 25 Working pitch circle diameter of roller gear, mm 77.2 77.2 Roller gear width, mm 36 36 Roller gear rotational angle for contact ratio calculation s (deg) 39"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003090_icelmach.2016.7732872-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003090_icelmach.2016.7732872-Figure6-1.png",
        "caption": "Fig. 6 Analysis model (CSVFRMDW)",
        "texts": [],
        "surrounding_texts": [
            "In this section, the operating principle of the CSVFRMDW is described. The fundamental frequencies of the magnetomotive force generated by the DC current Ff and permeance of the rotor Pf are expressed in (1) and (2), respectively. \u03b8ff NFF sin0= (1) ( )rrf NPPP \u03b8\u03b8 ++= sin10 (2) where F0 is the amplitude of the magnetomotive force, Nf is the number of pole pairs due to the DC currents, \u03b8 is the angular position of the rotor, P0 is the average permeance, P1 is the amplitude of the permeance, Nr is the number of rotor salient poles, and \u03b8r is the rotation angle of the rotor. The magnetic flux in the air gap \u03a6f is expressed by the product of the magnetomotive force and permeance as shown in (3). ( ){ } ( ){ }++\u2212 +\u2212 + = rrfr rrfr ff NNN NNN PF NPF \u03b8\u03b8 \u03b8\u03b8 \u03b8 cos cos 2 1 sin 10 00 (3) When either Nr Nf or Nr + Nf is equal to the number of pole pairs of the rotating magnetic field due to the AC current, the two synchronize and the rotor can rotate. Therefore, equations (4) and (5) must be satisfied. fra NNN \u00b1= (4) aarr NN \u03b8\u03b8 = (5) where \u03b8a is the rotation angle of the rotating magnetic field. In this paper, Na = 4, Nr =10 and Nf =6 are used. For other examples of the combination, Na =4, Nr = 7 and Nf = 3, and Na =5, Nr = 8 and Nf = 3 can be used."
        ]
    },
    {
        "image_filename": "designv11_64_0002457_978-3-319-08338-4_96-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002457_978-3-319-08338-4_96-Figure3-1.png",
        "caption": "Fig. 3 Basic pCCP problem. a Two clothoids composition. b Two clothoids and inter circular arc",
        "texts": [
            " In this work, we make a focus on the path generation for a car-like vehicle in a road environment. Problem: From Pi (xi , yi , \u03b8i ,\u03bai ) to Pf ( x f , y f , \u03b8 f ,\u03ba f ) passing through a point Pc at its maximal curvature \u03bamax , with \u03bai = 0 and \u03ba f = 0, find the minimum number of clothoids which satisfy the both configurations with curvature continuity along the path. Proposition 1 If the both end poses could not be accessed by clothoidal parametric variation, straight line or circular arc segments are added with the conformity on the orientation G1 and curvature continuity constraint G2. Figure 3a depicts the defined problem. The initial pose Pi and the final pose Pf are denoted by red arrows and the extension lines are displayed from each pose as i and f respectively.While fulfilling the pose configuration at both ends, clothoid segment C1 and C2 (\u2200s \u2208 [s0, sl ]) are generated from the point pc with same orientation and curvature to meet i and f at Pi and Pf respectively, where the inner-tangential circle rc with a radius \u03c1c and the center cc. For above case of \u03c0 2 < \u2223 \u2223\u03b8i \u2212 \u03b8 f \u2223 \u2223 < \u03c0, the minimal number of clothoid is two as proven in [11]",
            " Then it turns back to its middle center where former part is represented by the first clothoid and later part is expressed by the second clothoid. Another importance in fixing pc corresponds to when a vehicle path includes some region of constant angle of steering which is drawn as an circular arc segment. As usual, this steering behavior is shown at the region of highly curvatured road geometry section such as road intersection. In this case, the resultant solution only only two clothoids withot any circular arc segment could contain large differences with regard to the original path. Figure 3b shows the case where an circular arc segment is included between two clothoids. This solution is imposed in the cases where the both end positions are too close or/and the maximal curvature constraint is reached. Note that with the same configurations, the maximum curvature as given in Fig. 3a is greater then the case in b while it explicitly true if \u03c1c < \u03c1\u2032 c. This case is mainly useful for reconstructing the origianl reference path having regions of constant curvature as dealt with in the demonstrative example given in 3. The proposedmethodology to resolve the presented problem is firstly to determine pc with i and f from Pi and Pf . The determination of pc is described in the procedure given in Algorithm 2. By adjusting \u03b11 and \u03b41, C1 is generated until its end point C1(s0) to meet i for the distance to converge within a small constant threshold value defined by designer as De 1 < \u03b5 ( \u03b5 is a very small positive contant). After the convergence of C1 to i , C2 is generated with its resultant \u03ba2 and \u03b42 on G1 and G2 to check C2(s0) to be close to f for De 2 in \u03b5. In the non-converging case by clothoid parameter variation, arc segment is added as depicted in Fig. 3b. In order for C1 and C2 to converge to each configuration, the circular arc segment p\u0302c pc1, p\u0302c pc2 are included to make Cr by controlling its circular arc angle \u03b8c1, \u03b8c2 where Cr = \u03c1\u2032 c(\u03b8c1 + \u03b8c2) and \u03b8c2 is determined as the same way as \u03b8c1. Until the convergence of both segments to end configurations, the iterative variation on the parameters are performed for \u03b11, \u03b41, \u03b8c1 and \u03b8c2. The detailed procedure for finding clothoids parameters, taking into account the two cases given in Fig. 3, is described in Algorithm 1. In Algorithm 1, initial sharpness \u03b11i and deflection \u03b41i for C1 are assumed given by designer before entering the loop. It is noted that \u03b41i is set in the proposed work to (\u03b8i \u2212\u03b8 f ) 2 from the fact of \u03b41 + \u03b42 + \u03b8c = \u03b8i \u2212 \u03b8 f (G1 continuity at pc) with initially \u03b8c = 0, \u03b41 = \u03b42. The \u03b11i is set for C1(s0) to be located between i and pc, i.e. if C1(s0) resulting from \u03b41i and \u03bac = 1 \u03c1c is not located in i and pc, \u03b11i is doubled until the convergence conditions (line 2, 4 in Algorithm 1) is satisfied",
            " , n) m Arc boundary points 7: Elimination of redundant points 8: Boundary points determination 9: I (u\u0304, v\u0304) \u2212\u2192 P(x\u0304, y\u0304) Inverse mapping on Cartesian map 10: end procedure 11: procedure Local Planner pCCP((Pi , P f , pc)) 12: Configurations definition \u03b8i, f from Si and pc from Cj 13: Local planning using Algorithm 1 Two clothoids and/or circular arc 14: end procedure 15: Post processing Connecting the acquired solutions 16: end procedure The application of HT [23] for line and arc segments are performed as shown in Fig. 4a. There are unnecessary information and noises in the arc results, thus clustering by k-means nearest neighbor method are proceeded [24] and the cluster centers are sorted by minimal value of the distance summation as defined by, min \u03bc\u0302 (i=p, j=q)\u2211 i = j \u2225 \u2225\u03c3i \u2212 \u03c3 j \u2225 \u2225 , (10) where \u03c3i is the ith circle center in a cluster \u03bc\u0302 and L2(Euclidean) distance \u2016 \u2016 between the path and the circle center of minimal radius is selected as pc as shown in Fig. 3. The integrated result in this procedure is shown in 4b where the line and clothoid/circular arc segments are noted by Si and Cj for i = 1, 2, . . . , m, j = 1, 2, . . . , n where m, n are the number of the isolated segments respectively. From the results, isolation of straight and curved sections gives each problem for local planner pCCP, thus global continuous curvature trajectory is generated by connecting each solution result. Each solution has its own sharpness which is found from the geometric boundary configuration"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001967_000392184-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001967_000392184-Figure1-1.png",
        "caption": "Fig. 1. Supporting forces, measured at the feet of a healthy subject.",
        "texts": [],
        "surrounding_texts": [
            "Medicine and Sport, vol. 6: Biomechanics II, pp. 266-271 (Karger, Basel 1971)\nSupporting Forces Related to Overload Damage of Leg Amputees\nH. KRUSE, W. BAUMANN, and H. GROH\nInstitut flir Biomechanik, Deutsche Sporthochschule, Koln-Mlingersdorf\nWe tried to fix the form of load and its intensity on amputees by recording the rolling effects of feet and measuring their supporting forces. The experiment covered 70 test persons: 30 normal subjects, 20 shank amputees, and 20 thigh amputees.\nMethods of Measuring\nThe supporting force of a foot may be regarded as a measure for the load exerted on the movement apparatus. In order to measure the supporting forces we developed a stationary force plate which renders possible the recording of the chronological process of the supporting forces in both the vertical and the horizontal direction.\nResults of Measuring\nThe total of horizontal and vertical propulsing forces is indicated as support of the foot, the result being illustrated in figures 1, 2 and 3.\nSupporting Force\nIn the evaluation, the force values A-E measured in the 70 subjects are related to the respective weights ( = 100 Ofo). Of those standard values we formed the mean x, the standard deviation s, and the ratio of standard deviation to the mean\nv= mean mean\nD ow\nnl oa\nde d\nby :\nU ni\nve rs\nit\u00e9 d\ne P\nar is\n19 3.\n51 .8\n5. 19\n7 -\n1/ 23\n/2 02\n0 10\n:5 5:\n26 A\nM",
            "Supporting Forces Related to Overload Damage of Leg Amputees\n267\nFigure 4 shows 5 maxima and minima A-E compared with the body weights (= 100 0/0) of the 3 examined groups. The hatching indicates the scatter. Figure 5 shows the divergence of the 3 examined groups expressed as a percentage; force values of normal subjects being 1000/0. The hatching marks the results of statistical significance.\nCompared with healthy subjects, when the foot pushes off, the unload ing of the sound leg as well as of the prosthesis is, on an average, 9 to 100/0 smaller. This means a certain increase in the load.\nThe vertical pushing-off force B of the ball of the foot of the prosthe sis leg is 6-7 Ofo smaller than that in healthy subjects. The horizontal\nD ow\nnl oa\nde d\nby :\nU ni\nve rs\nit\u00e9 d\ne P\nar is\n19 3.\n51 .8\n5. 19\n7 -\n1/ 23\n/2 02\n0 10\n:5 5:\n26 A\nM",
            "KRusEIBAUMANN/GRoH\n268\npushing-off force in the walking direction in the case of the amputees is clearly smaller than in healthy people, viz. 20 Ofo less in the case of the sound leg, and 51 and 36% respectively in the prosthesis. The surplus load C of the amputee in the pushing-off phase is only of short duration and always stays below the maximum values (A, B) of normal subjects. The reduction of the pushing-off forces Band E clearly points to a smaller load in the gait of an amputee.\nD ow\nnl oa\nde d\nby :\nU ni\nve rs\nit\u00e9 d\ne P\nar is\n19 3.\n51 .8\n5. 19\n7 -\n1/ 23\n/2 02\n0 10\n:5 5:\n26 A\nM"
        ]
    },
    {
        "image_filename": "designv11_64_0000373_978-3-642-37694-8-Figure5.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000373_978-3-642-37694-8-Figure5.1-1.png",
        "caption": "Fig. 5.1 Cyclic Polling System",
        "texts": [
            " Once it flags an alert, its state does not change, even if additional triggering events were to occur, until the flag is reset by a loitering UAV. Hence there is at most only one alert waiting at an alert site. So, the perimeter patrol problem as envisaged here, constitutes a multi-queue, single-server, unit-buffer queueing system with deterministic (inter station) travel and service times. Since the UAV is constantly on patrol or is servicing a triggered UGS, the framework considered here is analogous to the cyclic polling system shown in Fig. 5.1. The basic model of a cyclic polling system consists of separate queues with independent Poisson arrivals (say, at rate \u03b1) served by a single server in cyclic order. A thorough analysis of such systems, and the various applications thereof, can be found in [39]. In the queueing literature, the performance measures of interest have been primarily the average queue length and average customer waiting time for service. Significant effort has been expended in evaluating these two metrics for polling systems under varying assumptions on the 5 Approximate Dynamic Programming Applied to UAV Perimeter Patrol 121 service time statistics, buffer capacity and service discipline [37, 38]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.4-1.png",
        "caption": "Figure 9.4 PM three-phase synchronous machines with single-layer and double-layer toothcoil windings: Qs= 6, p= 2, and q= 0.5.",
        "texts": [
            " In a tooth-coil P 3 Usph EPMph \u03c9sLd sin \u03b4s U2 sph Ld Lq 2 \u03c9sLd Lq sin 2\u03b4s (9.1) machine, the number of slots per pole and phase varies between q\u2208 [0.25, 0.5]. In particular, the machines are intended for low-speed direct drive machine applications where maximum torque is needed in a smaller volume. Tooth coils have smaller end windings and can work with a smaller stator yoke, so the architecture accommodates a larger maximum rotor diameter and length relative to the outer dimensions of the machine. Figure 9.4 gives cross-sectional views illustrating tooth-coil PM synchronous machines. The best tooth-coil machines produce a completely sinusoidal voltage, and therefore smooth torque can be achieved with sinusoidal stator currents. The control of this kind of a machine does not differ in principle from the control of any other ordinary rotating-field PM machine. Only the machine parameters are of interest. In principle, tooth-coil machines can also produce some reluctance torque. However, the proportion normally stays low, and the machines behave mostly like rotor surface magnet nonsalient-pole PM machines"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.13-1.png",
        "caption": "FIGURE 6.13",
        "texts": [
            " The choice of whether to combine them with the road spring and damper forces is entirely one of modelling convenience; the authors generally find the ease of debugging and auditing the model is worth the carriage of two not strictly necessary additional force generating terms. For the simplified modelling approach used in the lumped mass and swing arm models the road springs cannot be directly installed in the vehicle model as with the linkage model. Consider the lumped mass model when compared with the linkage model as shown in Figure 6.13. Clearly there is a mechanical advantage effect in the linkage model that is not present in the lumped mass vehicle model. At a given roll angle for the lumped mass model the displacement and hence the force in the spring will be too large when compared with the corresponding situation in the linkage model. For the swing arm model the instant centre about which the suspension pivots is often on the other side of the vehicle. In this case the displacement in the spring is approximately the same as at the wheel and a similar problem occurs as with the lumped mass model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002655_978-3-319-06590-8_87-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002655_978-3-319-06590-8_87-Figure1-1.png",
        "caption": "Fig. 1 Parameters of the linear perturbation",
        "texts": [
            " The nondimensional extended and generalized Reynolds equation is given by: @ @u H3 12g PKx @P @u \u00fe @ @ z H3 12g PKz @P @ z \u00bc 1 2 @ @u \u00f0q fCH\u00de \u00fe @ @s q H\u00f0 \u00de \u00f01\u00de with u \u00bc x r ; y \u00bc y CR ; z \u00bc z r ; H \u00bc h CR ; q \u00bc q q0 ; s \u00bc xst; g p \u00bc gp g0 \u00f02\u00de Herein \u03b7p is local radial averaged viscosity and fc is a viscosity factor. Both are determined according to [13]. Film pressure, film thickness and its time derivative can be written as the linear combination of steady state and perturbed part. The latter one does not have a stationary part: P\u00f0u; z\u00de \u00bc Pstat\u00f0u; z\u00de \u00fe @P @ X stat X \u00fe @P @ Y stat Y \u00fe @P @ X 0 stat X 0 \u00fe @P @ Y 0 stat Y 0 H\u00f0u; z\u00de \u00bc Hstat\u00f0u; z\u00de X sinu Y cosu; @H @s \u00f0u; z\u00de \u00bc X 0 sinu Y 0 cosu: \u00f03\u00de The dimensionful parameters of the linear perturbation are illustrated in Fig. 1. Integrating the expression defined by (3) into the generalized Reynolds Eq. (1) and neglecting all terms of second or higher order provides: @ @u H3 stat 12g pKx @ @u @P @ X stat \" # \u00fe @ @ z H3 stat 12g pKz @ @ z @P @ X stat \" # . . . \u00bc 1 2 @ @u q fC sinu\u00f0 \u00de \u00fe @ @u 3H2 stat sinu 12g pKx @Pstat @u \" # \u00fe @ @ z 3H2 stat sinu 12g pKz @Pstat @ z \" # ; \u00f04\u00de @ @u H3 stat 12g pKx @ @u @P @ Y stat \" # \u00fe @ @ z H3 stat 12g pKz @ @ z @P @ Y stat \" # . . . \u00bc 1 2 @ @u q fC cosu\u00f0 \u00de \u00fe @ @u 3H2 stat cosu 12g pKx @Pstat @u \" # \u00fe @ @ z 3H2 stat cosu 12g pKz @Pstat @ z \" # ; \u00f05\u00de @ @u H3 stat 12g pKx @ @u @P @ X 0 stat \" # \u00fe @ @ z H3 stat 12g pKz @ @ z @P @ X 0 stat \" # \u00bc q sinu; \u00f06\u00de @ @u H3 stat 12g pKx @ @u @P @ Y 0 stat \" # \u00fe @ @ z H3 stat 12g pKz @ @ z @P @ Y 0 stat \" # \u00bc q cosu: \u00f07\u00de For the iterative solution of (4)\u2013(7) all implicit dependencies of the other parameters and boundary conditions on the pressure have to be linearized as well"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002021_imccc.2015.215-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002021_imccc.2015.215-Figure1-1.png",
        "caption": "Figure 1. The rotation and force analysis of the quadrotor",
        "texts": [
            " For this part of the design, the implementation of the traditional way is through PID, fuzzy PID method, etc. In this paper we mainly present a new approach using the variable structure of modern control method to model and design the system. II. THE BASIC MODEL OF THE FOUR-ROTOR AIRCRAFT Four rotor aircraft flies through the lift four propeller produce, the principle is similar to the helicopter. Four rotor sits in front of a geometric symmetry cross bracket. Rotor wing, which relies on the change of each motor speed to achieve flight attitude control, is controlled by the motor. In the figure 1, Front-end rotor 1 and back-end rotor 3 counterclockwise, rotor 2 on the left side and rotor 4 on the right side clockwise. This rotation is to Balance the rotor rotation of the torsion moment. When hovering, the rotate speed of four rotor should be equal to offset the torsion moment. Amount at the same time to increase or decrease the speed of four rotor, will cause rising or falling movement. Increase the revolving speed of a rotor, while an equal amount to reduce the speed of the other rotor of the same group, to produce pitch and roll motion"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000866_iros.2014.6943044-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000866_iros.2014.6943044-Figure3-1.png",
        "caption": "Fig. 3. Relation between Coordinate Systems \u03a3i and \u03a3k",
        "texts": [
            " Note that the velocity information of each caster unit is shared in a centralized control system as shown in the simulation and the experiment later. However, the proposed method is also applicable to a distributed control system in which each caster unit independently calculates its own velocity and communicates it with each other. Next, we denote the position of the origin of the coordinate system \u03a3i relative to that of \u03a3k by kri = [krix kriy ] T , and the orientation of the coordinate system \u03a3i relative to that of \u03a3k by k\u03b8iz , as shown in Fig. 3. In terms of krix, kriy , and k\u03b8iz , the relationship between cx\u0307i, cy\u0307i, c\u03b1\u0307i and cx\u0307k, cy\u0307k, c\u03b1\u0307k is expressed by   cx\u0307k cy\u0307k c\u03b1\u0307k   =   cos k\u03b8iz \u2212 sin k\u03b8iz 0 sin k\u03b8iz cos k\u03b8iz 0 0 0 1     cx\u0307i + c\u03b1\u0307i kriy cy\u0307i \u2212 c\u03b1\u0307i krix c\u03b1\u0307i   . (2) We define the state vector \u01eb as the pose of the attachment point in coordinate system \u03a3i. The pose comprises the position of the attachment point of the i-th caster unit kri and the orientation k\u03b8i: \u01eb = [ krix kriy k\u03b8iz ]T . (3) Since the relationships among velocities of the caster units and \u01eb are non-linear, \u01eb cannot be directly calculated from Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure2.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure2.6-1.png",
        "caption": "Fig. 2.6 (a) Lateral contraction and (b) lateral expansion of solid bodies subjected to an axial force",
        "texts": [
            " In the left case, an external load F is acting on the side surfaces of the blocks, such that a shear force F is induced on the contact area. The right case contains three bodies with the side load of F/2 applied to the upper and lower blocks and the side load F in another direction. The load F on the middle block is distributed as two shearing forces over both the upper and lower interfaces, each with a shearing force of F/2. If a solid body is subjected to axial tension, it contracts laterally at the same time. On the other hand, an object with an axial compression would have increases in its lateral side lengths as described in Fig. 2.6. This phenomenon is described by the material property called Poisson\u2019s ratio. It is the ratio of lateral contraction (extension) strain (\u03b5lateral) to axial extension (contraction) strain (\u03b5axial) in the direction of direct tensile (compressive) force. It is denoted by \u03bd (nu) and is represented as 2.2 Solid Properties 41 \u03bd \u00bc \u03b5lateral \u03b5axial \u00f02:16\u00de The equation of Poisson\u2019s ratio contains a negative sign to keep normal materials with a positive ratio because the signs of the lateral strain and axial strain are opposite to each other"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.14-1.png",
        "caption": "Fig. 6.14 Configuration of an injection molding machine",
        "texts": [
            " The last step is the finishing of the formed product into the final product. 156 6 Common Manufacturing Process Injection molding is a widely used manufacturing process mainly used for the mass fabrication of plastics as well as biopolymers. High versatility of the process makes it highly adaptable to different factors, including size, shape, complexity, and components of the products. Other advantages of this process include its high versatility, rapid production rate, low production cost, and low scrap rate. An injection molding machine (Fig. 6.14) can produce tens of thousands of parts successively in a short period, leading to an effective production rate. Even though the initial cost of the process is very expensive, especially the molding dies, the per-part cost is extremely low since production cost tends to drop greatly as more parts are created. Besides, injection molding produces lower scrap amounts than older manufacturing processes like machining. There are four basic components of the mold that are important in injection modeling: sprue, gates, runners, and vents"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000218_978-3-319-52219-7-Figure3.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000218_978-3-319-52219-7-Figure3.1-1.png",
        "caption": "Fig. 3.1 Typical simulation method for metamaterials using unit cells. Unit cell showing electric/magnetic boundary conditions used in the Microwave Studio [6] simulations. Periodic boundary conditions are applied along the x-axis and y-axis to emulate the array of metamaterials in the X-Y plane. The direction of propagation of the electromagnetic field is along the z-axis, the electric field is oriented along the y-axis, the magnetic field along the x-axis",
        "texts": [
            " If a mirror symmetry exists in the metamaterial structure that is being simulated, two oppositely located electric and/or magnetic boundary conditions maybe defined to emulate the entire array thus saving computation time and memory. The reason for this is that the boundary conditions work like mirrors and, when two mirrors are placed in front of each other, infinitely many mirror images of the unit structure are created. The designer has to ensure that in addition to geometric symmetry, field symmetry is also matched by the boundary condition. The boundary condition of a metamaterial unit cell is shown in Fig. 3.1, where the unit cell is part of a metamaterial array in the X-Y plane. The excitation of the cell is provided by a linearly y-polarized electromagnetic wave propagating 3.2 Design for Fabrication in Foundry Processes 43 in the z direction. In order to mirror the unit cell, the two walls perpendicular to the y axis are defined as electric boundaries and the walls perpendicular to x axis are defined as magnetic boundaries. The top and bottom walls are defined as open boundaries for the EM wave propagation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001325_iedec.2014.6784673-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001325_iedec.2014.6784673-Figure4-1.png",
        "caption": "Figure 4: CAD drawing of typical design",
        "texts": [
            " The dimensions of the casing and the available motor are defined, so it is known: Using the above specifications the students were required to perform calculations for the required power, a blockage factor due to the presence of the blades, the optimal leading edge angle taking into account the blockage factor, the trailing edge angle, with the blockage factor included, the volute coupling, the slip factor, the velocity match and an estimate of the number of blades to be used. Following on from the above calculations the students produced individually a CAD drawing with a typical one shown on Figure 4. Groups of 5 students were then formed and in a committee like atmosphere they debated and chose the best design within each group. The best design for each group was manufactured using a 3-D printer [9]. This builds the impeller layer by layer from composite material powder as shown on Figure 5. Each layer thickness was 0.1 mm and the vertical built speed was 23 mm/hr. This is followed by two further stages called processing and chemical curing. First excessive powder was collected using a vacuum cleaner for recycling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000921_amm.628.283-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000921_amm.628.283-Figure3-1.png",
        "caption": "Fig. 3. Circular tool modes analysis mode shapes",
        "texts": [
            " The analysis frequency is set range 10000Hz to 30000Hz and according to the actual needs, prestress should be applied in the cutting edge,along the X , Y , Z three directions. In this paper the applied prestress is Fx=9N,Fy=15N,Fz=-14N, each modes frequency is shown in table 2. Tab. 2. Ten modes frequency Modes 1 2 3 4 5 6 7 8 9 10 Frequency\uff08Hz\uff09 20511 20511 23191 23191 24615 24617 27069 28839 28840 34680 Based on the frequency distribution of ten modes analysis, three modes is extracted and analyzed, as is shown in figure 3, they are second modes, ninth modes and seventh modes. The second modes analysis results is more relaxed, the ninth is more intense and above of them the maximum amplitude distribute in the outermost of the tool, but those modes vibration direction are inconsistent. The seventh analysis results is much more relaxed, maximum amplitude uniformly distribute in outer edge of the circular tool and the direction of seventh vibration modes is consistent. It is an ideal modes and the frequency is 27069Hz"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003196_cistem.2014.7076970-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003196_cistem.2014.7076970-Figure2-1.png",
        "caption": "Fig. 2 .Structure de l'inducteur",
        "texts": [
            " Une analyse en trois dimensions, du champ magn\u00e9tique, est faite (en utilisant un logiciel d'\u00e9l\u00e9ments finis 3D) pour \u00e9tudier les performances de l'inducteur propos\u00e9. L'effet de la culasse du stator sur la r\u00e9partition du champ magn\u00e9tique est \u00e9galement \u00e9tudi\u00e9. Un test de l\u2019inducteur \u00e0 l\u2019azote liquide est effectu\u00e9 pour valider le principe de cette structure et relever la carte du champ magn\u00e9tique pour la comparer avec le calcul num\u00e9rique. II. STRUCTURE DU MACHINE Apr\u00e8s avoir pr\u00e9sent\u00e9 le principe de notre inducteur, nous d\u00e9crirons les m\u00e9thodes de calculs utilis\u00e9es La Fig. 2 montre la structure de l'inducteur propos\u00e9. Il est compos\u00e9 de deux bobines supraconductrices coaxiales. Les bobines sont aliment\u00e9es par des courants \u00e9lectriques ayant la m\u00eame direction qui g\u00e9n\u00e8rent un champ magn\u00e9tique (B1, B2). Un supraconducteur massif est plac\u00e9 entre les deux bobines. Le supraconducteur est inclin\u00e9 le long de la longueur de cet inducteur. Ce supraconducteur est utilis\u00e9 comme une barri\u00e8re magn\u00e9tique et est situ\u00e9 entre les deux sol\u00e9no\u00efdes, cr\u00e9ant ainsi un champ magn\u00e9tique variable tel que mentionn\u00e9 dans la Fig. 1. Afin d'am\u00e9liorer la distribution de l'induction radiale, on utilise aussi un mat\u00e9riau ferromagn\u00e9tique entre les deux bobines. Le supraconducteur est ins\u00e9r\u00e9 entre les pi\u00e8ces ferromagn\u00e9tiques comme repr\u00e9sent\u00e9 sur la Fig. 2. Le fer guidera le champ magn\u00e9tique vers le centre de l'inducteur et donc r\u00e9duira le champ de fuite \u00e0 proximit\u00e9 des bobines. Pour simplifier le syst\u00e8me de refroidissement de cette machine, nous avons choisi d\u2019avoir un induit tournant et un inducteur supraconducteur fixe et immerg\u00e9 dans un liquide cryog\u00e9nique comme il est montr\u00e9 dans la Fig. 3. La structure d'inducteur propos\u00e9e est en mesure de fournir une machine \u00e0 deux p\u00f4les et peut \u00eatre assimil\u00e9 aux machines synchrones \u00e0 griffes. Cette machine dispose d'un couple \u00e9lectromagn\u00e9tique \u00e9lev\u00e9e en raison de l'augmentation de la densit\u00e9 de flux dans l'entrefer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000482_20140824-6-za-1003.01497-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000482_20140824-6-za-1003.01497-Figure2-1.png",
        "caption": "Fig. 2. Schematic of the location of the thruster force vectors, moment arms, center of gravity and buoyancy and the axis system, where the pitch angle is \u221210o.",
        "texts": [
            ", performing complex manoeuvres or if the system unexpectedly changes due to loss or gain of buoyancy. Moreover, in operation there is an engineering need to place magnitude and rate constraints on the control signals for more demanding missions. Such constraints are not included in the PID design, leading to a control law that cannot be applied or degraded/unacceptbale performance. These reasons justify the use of designs, such as MPC, where such constraints can be imposed in the design stage. The AUV can be modeled as two coupled second-order systems; depth and pitch, see Figure 2, and the equations of motion are q\u0307v =\u2212 1 Iy [xTvfTvf + xTvrTvr \u2212 zgW sin \u03b8 + 1 2 \u03c1V 2/3CDq|qv|qv], qv = \u222b t 0 q\u0307v dt, \u03b8 = \u222b t 0 qv dt, (1) for pitch and for depth w\u0307v = 1 mz [Tvf cos \u03b8 + Tvr cos \u03b8 + (W \u2212B) \u2212 1 2 \u03c1V 2/3CDw|wv|wv], wv = \u222b t 0 w\u0307v dt, z = \u222b t 0 wv dt. (2) The outputs from the controller are the force demands for each thruster (in Newtons (N)). To translate these force values into thruster speed set-points, that can be sent to the thruster controller, the inverse of the following thrust equation is used n = 60\u00d7 [ Tdemand \u03c1KTD4 ]0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-FigureA.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-FigureA.8-1.png",
        "caption": "Fig. A.8 Sign of the second derivative of a curve: a x\u2013y plane: a curve which is bending away from normal vector n gives a negative sign (left) and a curve which is bending towards the normal vector n gives a positive sign (right); b x\u2013z plane: a curve which is bending away from normal vector n gives a negative sign (left) and a curve which is bending towards the normal vector n gives a positive sign (right)",
        "texts": [
            "165) In the scope of the finite element method, the coefficient matrix A is equal to the global stiffness matrix K . Since this matrix is symmetric and may have a special triangular structure, more sophisticated solution procedures can be introduced. 406 Appendix A: Mathematics \u2022 Slope-intercept form: Given is a slope m and the y-intercept b: y = b + m \u00d7 x . (A.170) \u2022 Two-point form: Given are the points (x1, y1) and (x2, y2): y \u2212 y1 x \u2212 x1 = y1 \u2212 y2 x1 \u2212 x2 , (A.171) y = ( y1 \u2212 y1 \u2212 y2 x1 \u2212 x2 x1 ) + y1 \u2212 y2 x1 \u2212 x2 \u00d7 x . (A.172) A.14.2 Sign of Second Derivative of a Curve See Fig.A.8. Appendix B Mechanics B.1 Centroids The coordinates (zS, yS) of the centroid S of the plane surface shown in Fig.B.1 can be expressed as zS = \u222b z dA \u222b dA , (B.1) yS = \u222b y dA \u222b dA , (B.2) where the integrals \u222b z dA and \u222b y dA are known as the first moments of area.4 In the case of surfaces composed of n simple shapes, the integrals can be replaced by summations to obtain: zS = \u2211n i=1 zi Ai\u2211n i=1 Ai , (B.3) yS = \u2211n i=1 yi Ai\u2211n i=1 Ai . (B.4) 4A better expression would be moment of surface since area means strictly speaking the measure of the size of the surface which is different to the surface itself"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001102_rev.2014.6784184-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001102_rev.2014.6784184-Figure4-1.png",
        "caption": "Figure 4. Surfboard positioning and axis definitions.",
        "texts": [
            " A geometrical reference point is positioned 18 cm far from the first sensor on the x axis and, the last sensor is positioned 90 cm from the reference, giving 90 cm for the x axis. The y axis size is 19 cm. To get the surfboard\u2019s positioning in relation to the gravity axis an IMU was mounted on the electronic board. The IMU has an accelerometer and a gyroscope providing information about the 3-axis acceleration components and two rotation angles (roll around the x-axis and pitch around the y-axis), Fig. 4. Data from the sensors is acquired and processed using an ATEMEGA1280 microcontroller. This component has embedded an analog to digital converter (ADC) among other peripherals like real time clock circuit, Universal Asynchronous Receiver/Transmitter (UART) communication, etc. In order to manage data acquisition, an analog multiplexer circuit was used for which a cascade circuit combined 24 sensor inputs. When one of the inputs is connected to the output of the cascade multiplexer circuit, it becomes connected to the ADC circuit through a Wheatstone bridge, Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000586_ascc.2015.7244376-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000586_ascc.2015.7244376-Figure2-1.png",
        "caption": "Figure 2: Quarter-car model",
        "texts": [
            "1802, respectively. The simulation results show that the proposed controller effectively reduces the sideslip angle and yaw angle in yaw and roll flight condition, and can resist the influence of air disturbances. B. Application to 1/4 Active Suspensions 1) 2-DOF Quarter-Car Model and Control Problem Formation: 2-DOF quarter-car models are widely used in suspension analysis and design, because they capture major characteristics of a real suspension system. A generalized quarter-car suspension model is shown in Figure 2, where ms and mu stand for the sprung mass and unsprung mass; ks and cs are stiffness and damping of the suspension system, respectively. In addition, (ks ,cs ) consists of the socalled passive suspension; ku represents the tire stiffness; xs \u2212 xu denotes the suspension stroke and xg is vertical ground displacement caused by road unevenness. Moreover, u f is the active control force provided by a hydraulic actuator. The model parameters are given in Table 1 for the controller design. Based on this suspension model, the linearized dynamic equations of the sprung and unsprung mass can be established [6]: ms x\u0308s +ks (xs \u2212xu)+ xs (x\u0307s \u2212 x\u0307s ) = u f (18a) mu x\u0308s \u2212ks (xs \u2212xu)\u2212xs (x\u0307s \u2212 x\u0307s )+ku(xu \u2212 xg ) = u f (18b) Define a set of state variables x1 = xs \u2212 xu , x2 = x\u0307s , x3 = xu \u2212xg , x4 = x\u0307u , the state description of the car motion can be obtained as x\u0307(t ) = \u23a1 \u23a2\u23a2\u23a2\u23a3 0 1 0 \u22121 \u2212 ks ms \u2212 cs ms 0 cs ms 0 0 0 \u22121 ks mu cs mu \u2212 ku mu \u2212 cs mu \u23a4 \u23a5\u23a5\u23a5\u23a6x(t ) + \u23a1 \u23a2\u23a2\u23a3 0 0 \u22121 0 \u23a4 \u23a5\u23a5\u23a6w(t )+ \u23a1 \u23a2\u23a2\u23a2\u23a3 0 us ms 0 \u2212 us mu \u23a4 \u23a5\u23a5\u23a5\u23a6u(t ) , (19) where w = x\u0307g denotes the disturbance input caused by road roughness, us is the maximum active force and u = u f /us is the normoalized active force"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001446_j.proeng.2014.06.045-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001446_j.proeng.2014.06.045-Figure1-1.png",
        "caption": "Fig. 1. Frame RJ (x,y,z) with its origin fixed at the universal joint and mast frame Rm (x\u2019,y\u2019,z\u2019)",
        "texts": [
            " The direction of the measured force is parallel to the wishbone plane. The pumping is a movement in which the athlete transmits a force Fa (supposed parallel to the fore-arms) to the rig to move the board during a time t. Thus, a tilt sensor is needed to measure the angle between the forearm and the displacement of the force origin. The mechanical power P developed during this time is given by t d. t W P lFa (1) where W is the work when the force Fa acts during a time t resulting in a displacement L =AB of its point of application (See Fig. 1). W is expressed by lFa d.W (2) In practice, the discrete version of the integral in relation (2) is the sum, thus n i W 1 il.Fa (3) With li = being an elemental displacement of the force origin resulting from the work done. The displacement of the point of application of F can be calculated considering the well-known direct geometric model in robotics. In this model, the point of application of F was taken as the origin of a frame referred to as boom frame Rb that can be deduced from mast frame Rm by a vector translation L2. The universal joint frame RJ with unit vectors i, j and k along x, y and z respectively, was chosen as shown in Fig. 1. The origin of RJ was located at the universal joint which connects the rig to the board. z-axis is along the mast while x-axis and y-axis are in the wishbone boom plane. Rm is deduced from RJ by a vector translation L1 and RJ is deduced by a rotation about zb of an angle (See Fig.2) from board reference frame RS (xS, yS, zS) in which zero calibration of the two IMU has been performed. The homogenous matrix T (Khalil and Dombre, 2004; 2007) representing in our case the transformation from RS to Rb is then defined as 1000 z y x pmCmCmCmSmS pmCmSmSmSmCmCmCmSmSmSmSmC pmSmSmCmSmCmSmCmCmSmSmCmC T (4) px = L2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002827_jae-162095-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002827_jae-162095-Figure1-1.png",
        "caption": "Fig. 1. Structure of the magnetic gears.",
        "texts": [
            " For example, the positioning accuracy of a magnetic gear is required when the magnetic gear is used in a part of the joint of a robot arm. So the positioning property of a cylindrical magnetic gear were studied in [6,7]. In this paper, torque transmitting property is investigated through the experiment and the simulation. Also the effect of number of stacks is concerned. \u2217Corresponding author: Yoshinori Ando, Division of Mechanical Science and Technology, Gunma University, 1-5-1 Tenjincho, Kiryu 376-8515, Japan. E-mail: ando@gunma-u.ac.jp. 1383-5416/16/$35.00 c\u00a9 2016 \u2013 IOS Press and the authors. All rights reserved Figure 1 shows the proposed magnetic gear with the multiple layers of the inner and the outer rotors. The ring magnets and the iron plates with gear-like shape are stacked in both rotors. These magnets are magnetized along the axial direction and the same poles face each other across the iron plates to concentrate the magnetic flux. The center steel segment is made of electromagnetic irons and nonmagnetic material. Figure 2 shows photos of parts of the proposed magnetic gear Type A. As shown in Fig. 2, the magnetic gear is structured of (a) an outer iron plate of Type A, (b) an outer ring magnet, (c) an inner iron plate of Type A, and (d) an inner ring magnet",
            " Gr is the speed ratio when the outer is fixed. In this study, Pi and Po are equal to 5 and 11 respectively. So Nc is calculated as 16 by Eq. (1). Also Gr is calculated as 3.2 by Eq. (1). The rotatory direction of the output is same to that of the input because the sign of Gr is plus. In many surface permanent magnetic gears research, magnets are magnetized by the radial direction. Inserting a long part axially with magnets is difficult to assemble. So a new proposed gear has a stackable structure as show in Fig. 1(b) to assemble easily. At this structure, magnets are magnetized by the axial direction. This type magnetic gear rotates by the magnetic flux which is concentrated by opposing same pole of magnets. It makes assembling easily because every layer is stacked after making each layer. And it becomes possible to adjust the maximum transmit torque by changing the number of stacks. Figure 3 shows an experimental device. The input (high speed) rotor and the output (low speed) rotor are connected to the motor and the powder brake respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.27-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.27-1.png",
        "caption": "Fig. 3.27 Simply supported beam: a single force case; b distributed load case",
        "texts": [
            "188) This finite element solution is equal to the analytical solution. \u2022 \u222b Nq(\u03be)d\u03be will be checked for the first component of f eI f1Z = 1\u222b \u22121 1 4 ( 2 \u2212 3\u03be + \u03be3 ) ( \u2212q0 4 ) (1 + \u03be) L 4 d\u03be. (3.189) one-point rule f1Z = \u2212q0L 16 , (3.190) two-point rule f1Z = \u22125q0L 144 , (3.191) three-point rule f1Z = \u22123q0L 80 . (3.192) 3.3 Finite Element Solution 139 3.5 Example: Simply supported beam problems\u2014comparison between finite element and analytical approach Given is a simply supported Euler\u2013Bernoulli beam as shown in Fig. 3.27. The length of the beam is L and the bending stiffness is EI . Consider two different load cases in the following: (a) a single force F acting in the middle of the beam. (b) A constant distributed load q. Calculate based on (I) one single and (II) four beam finite elements of equal length the deformations at the nodes and in the middle of each element. In addition evaluate the maximum stress at the nodes and in the middle of the element. Compare your results with the analytical (exact) solution and calculate the relative error"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure9-1.png",
        "caption": "Figure 9. (a) The system of equivalent springs without a static force for the system shown in Figure 8(a), and (b) the system of equivalent springs for the same system with a static force.",
        "texts": [
            " In addition, based on Figure 8(a), one can find the lengths of the springs in the static equilibrium position l1\u00bc ffiffiffiffiffi 13 p m, l2\u00bc 3m, l3\u00bc 5m and l4\u00bc 6m. The position of each of the spring is determined based on Figure 8(a) and (b) as \u20191\u00bc 3 / 2\u00feb, \u20192\u00bc 3 /2, \u20193\u00bc 3 /2\u2013g and \u20194\u00bc , where \u00bc arctan 2=3\u00f0 \u00de \u00bc 0:588 and \u00bc arctan 4=3\u00f0 \u00de \u00bc 0:927. The angle k from equation (1) is defined by tan 2 k \u00bc P4 i\u00bc1 ki 1 i li sin 2\u2019iP4 i\u00bc1 ki 1 i li cos 2\u2019i \u00bc 0:75 \u00f042\u00de which gives k\u00bc 0.322\u00bc 18.435o. Equation (2a,b) yields the stiffness coefficients of the equivalent system plotted in Figure 9(a) kI \u00bc X4 i\u00bc1 ki cos2 \u2019i k\u00f0 \u00de \u00fe i li sin2 \u2019i k\u00f0 \u00de \u00bc 3989:2 N=m, kII \u00bc X4 i\u00bc1 ki sin2 \u2019i k\u00f0 \u00de \u00fe i li cos2 \u2019i k\u00f0 \u00de \u00bc 5995:4 N=m \u00f043\u00de Note that this equivalent system does not contain the static force mg and the two new springs are not pre-stressed. The invariant related to the new stiffness coefficients given by equation (12) is satisfied. If one would like to establish the equivalent system with two springs in which the force mg is retained (Figure 9(b)), the facts expressed in Theorem 2 are at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from used. The angle k is the same as the one defined by equation (42), while equation (15a,b) for the assumed values of lI \u00bc ffiffiffiffiffi 40 p m, lII \u00bc ffiffiffiffiffi 10 p m gives k 0 I \u00bc kI mg cos k lII \u00bc 3719:2 N=m, k 0 II \u00bc kII mg sin k lI \u00bc 5950:4 N=m \u00f044a; b\u00de The corresponding static deflections of these springs are I \u00bc mg sin k k 0 I \u00bc 0:0765 m, II \u00bc mg cos k k 0 II \u00bc 0:143 m \u00f045a; b\u00de It is easy to verify that the invariant (19) from Corollary 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.13-1.png",
        "caption": "FIGURE 8.13",
        "texts": [
            " In CAD, more commonly employed joints (e.g., revolute, translation, cylindrical) have been replaced by assembly mates. Like joints, assembly mates remove degrees of freedom between parts. Each independent movement permitted by a constraint (either a joint or a mate) is a free degree of freedom. The free degrees of freedom that a constraint allows can be translational or rotational along the three perpendicular axes. For example, a concentric mate between the propeller assembly and the case of a single-piston engine shown in Figure 8.13 allows one translational DOF (movement along the center axisdin this case the X-axis) and one rotational DOF (rotating along X-axis). Since the case assembly is stationary, serving as the ground body, the propeller assembly has two free DOF. Adding a coincident mate between the two respective faces of the engine case and the propeller shown in Figure 8.13 removes the remaining translational DOF, yielding a desired assembly that resembles the physical situationdthat is, with only the rotational DOF (along the X-axis). In creating a motion model, instead of all movements being completely fixed, certain DOF (translational and/or rotational) are left to allow desired movement. Such a movement is either driven by a motor, resulting in a kinematic analysis, or determined by a force, leading to a dynamic analysis. For example, a rotary motor is created to drive the rotational DOF of the propeller in the engine example",
            " Essentially, 3D contact constraint applies a force to separate the parts when they are in contact and prevent them from penetrating each other. The 3D contact constraint becomes active as soon as the parts are in contact. As mentioned before, an unconstrained body in space has six degrees of freedom: three translational and three rotational. When mates are added to assemble parts, constraints are imposed to restrict the relative motion between them. Let us go back to the engine example shown in Figure 8.13. A concentric mate between the propeller and the engine case restricts movement on four DOF (Ty, Tz, Ry, and Rz) so that only two movements are allowed, one translational (Tx) and one rotational (Rx). To restrict the translational movement, a coincident mate is added. A coincident mate between two respective faces of the propeller and the engine case restricts movement of three DOF: Tx, Ry, and Rz. Even though combining these two mates achieves the desired rotational motion between the propeller and the case, redundant DOF are imposed: Ry and Rz in this case",
            " For a given motion model, the number of degrees of freedom can be determined using Gruebler\u2019s count, defined as D \u00bc 6M N O (8.105) where D is Gruebler\u2019s count representing the overall degrees of freedom of the mechanism. M is the number of bodies excluding the ground body. N is the number of DOF restricted by all mates. O is the number of motion drivers (motors) defined in the system. Consider a motion model consisting of the propeller, the engine case, and the rotary motor, in which the propeller is assembled to the engine case by a concentric and a coincident mate (see Figure 8.13). Gruebler\u2019s count of the two-body motion model is D \u00bc 6 1 \u00f04\u00fe 3\u00de 1 \u00bc 2 However, we know that the propeller can only rotate along the X-axis; therefore, there is only 1 DOF for the system (Rx), so the count should be 1. After adding the rotary motor, the count becomes 0. The calculation gives us 2 because there are two redundant DOF, Ry and Rz, which are restrained by both concentric and coincident mates. If we remove the redundant DOF, the count becomes D \u00bc 6 1 \u00f04\u00fe 3 2\u00de 1 \u00bc 0 Another example is a door assembled to a door frame by two hinge joints",
            " Most motion solvers detect redundancies and ignore redundant DOF in all but dynamic analyses. In dynamic analysis, the redundancies can lead to possibly incorrect reaction results, yet the motion is correct. For complete and accurate reaction forces, it is critical to eliminate redundancies from the mechanism. The challenge is to find the mates that impose nonredundant constraints and still allow the intended motion. A combination of a concentric and a coincident mate is kinematically equivalent to a revolute joint, as illustrated in Figure 8.13, between the propeller and the engine case. A revolute joint removes five DOF (with no redundancy); however, combining a concentric and a coincident mate removes seven DOF, among which two are redundant. Using assembly mates to create motion models almost guarantees redundant DOF. The best strategy is to create an assembly that closely resembles the physical mechanism by using mates that capture the characteristics of the motion revealed in the physical model. That is, an assembly should first be created that correctly captures the mechanism\u2019s kinematic behavior"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001724_ssd.2015.7348203-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001724_ssd.2015.7348203-Figure1-1.png",
        "caption": "Fig. 1. Cartesian Space",
        "texts": [
            " The idea of BCM is inspired by CVM and LCM. It is called the beam curvature method, because it takes the CVM method as a basis. This approach calculates the efficient heading that will be sent to Curvature Velocity Method to obtain the best translational and rotational velocities. The CVM approximates that all obstacles are circles and the robot moves along arc of a circle . The curvature is given by c = v/\u03c9 where v and \u03c9 are the translational and rotational velocities. CVM converts cartesian space (Fig.1) to configuration space (Fig.2) where the 3 dimensions are rotational velocity \u03c9, translational velocity v and curved distance to obstacle dc. The robot kinematics constraints [10] imposed by CVM are:\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 \u03c9 > \u03c9cur \u2212 \u03c9amaxT \u03c9 < \u03c9cur + \u03c9amaxT v < vcur \u2212 vamaxT (1) where \u03c9cur and vcur are the current angular and linear velocities, vamax and \u03c9amax are the maximum acceleration of linear and angular velocities and T is the decision time. In the first step, we should find the parameters (max curvature, min curvature and impact distance dc) of each obstacle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003391_andescon.2016.7836198-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003391_andescon.2016.7836198-Figure1-1.png",
        "caption": "Fig. 1. The translational manipulator with the corresponding experimental setup.",
        "texts": [],
        "surrounding_texts": [
            "The model of Lagrange given by equation (??) is used for a great class of nonlinear control systems. For the general case, let us assume that vectors q and u are of order m. The control objective of the backstepping control system as follows. Employing the measured vector q, designing a control law u capable of making that all elements of the output vector q follows arbitrary reference signals of the desired vector qd(t), correspondingly, meeting design specifications previously established. The application developed in this work uses the backstepping algorithm described in [?], which is briefly presented for the sake of clarity. Let us define the tracking error e as follows e = q\u2212 qd (4) In (??), qd is the vector of desired trajectories. Let us define z1 = e. Its derivative results z\u03071 = e\u0307 = q\u0307\u2212 q\u0307d = e\u03bd (5) Let us establish e\u03bd as the virtual control vector and select the following stabilizing function \u03b11 = \u2212Kz1 (6) In ??, K is a diagonal matrix of the for K = kI, where k is a positive constant, and I is the identity matrix. The state variable error can be defined as z2 = e\u03bd \u2212 \u03b11 = q\u0307\u2212 q\u0307d +Kz1 = q\u0307\u2212 q\u0307r (7) In (??), q\u0307r is defined as q\u0307r = q\u0307d \u2212Kz1 (8) Therefore, equation (??) can be rewritten as z\u03071 = z2 + \u03b11 = \u2212Kz1 + z2 (9) The derivative of the vector error z2 takes on the form z\u03072 = q\u0308\u2212 q\u0308d +Kz\u03071 (10) Employing equations (??) and (??), equation (??) can be formulated as z\u03072 = M\u22121(q)[u\u2212P(q, q\u0307)q\u0307\u2212d(q)]\u2212q\u0308d+Kz2\u2212K2 z1 (11) The stability analysis of the feedback control system uses the Lyapunov\u2019s direct method. Let us consider the following Lyapunov\u2019s function V = 1 2 zT1 K1z1 + 1 2 zT2 M(q)z2 + 1 2 \u02d9\u0303q T M(q) \u02d9\u0303q \u2261 1 2 xT1 P1x (12) In (??), \u02d9\u0303q = q\u0307 \u2212 \u02d9\u0302q is the speed estimation error, xT = [ zT1 zT2 \u02d9\u0303q T ] is the state vector, and P1 = diag [K1 M(q) M(q)] is a diagonal matrix. Taking the total derivative of V , we obtain V\u0307 = \u2212\u03b11\u2016z1\u20162 \u2212 \u03bb2\u2016z2\u20162 \u2212 \u03bb3\u2016 \u02d9\u0303q\u20162 (13) In (??), \u03b11, \u03bb2 and \u03bb3 are positive constants. Therefore V\u0307 is negative definite, which assure the asymptotic stability of the feedback control system. The stability analysis performed in [?] demonstrates that the following control law is able to stabilize the feedback control system u = M(q)q\u0308+P(q, \u02d9\u0302q)q\u0307r+d(q)\u2212Kd( \u02d9\u0302q\u2212 q\u0307r)\u2212K1 z1 (14) The speed observer to estimate \u02d9\u0302q takes on the form \u02d9\u0302q = q\u0307d + Ld(q\u2212 q\u0302) (15) The control gains Kd, K1 and Ld are diagonal definite positive matrices with Kd = kdI, K1 = k1I and Ld = `dI, where kd, k1 y `d are positive constants."
        ]
    },
    {
        "image_filename": "designv11_64_0003196_cistem.2014.7076970-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003196_cistem.2014.7076970-Figure5-1.png",
        "caption": "Fig. 5.Mod\u00e9lisation coulombienne d'un sol\u00e9no\u00efde",
        "texts": [
            " Param\u00e8tres g\u00e9om\u00e9triques Rayon ext\u00e9rieur Rex (m) 0,25 Rayon int\u00e9rieur Ri (m) 0,15 Distance entre les sol\u00e9no\u00efdes C (m) 0,15 Longueur du sol\u00e9no\u00efde L (m) 0,15 Epaisseur de l'\u00e9cran supraconducteur E (m) 0,01 Densit\u00e9 de courant J (A/mm2) 100 utilisons le logiciel COMSOL-multi-physique pour l'\u00e9tude de notre inducteur. La m\u00e9thode coulombienne est une m\u00e9thode sans courant, donc une charge magn\u00e9tique doit \u00eatre d\u00e9finie pour repr\u00e9senter les bobines. Pour chaque sol\u00e9no\u00efde, nous imposons l'aimantation \u00e9quivalente M, Fig. 5, selon les \u00e9quations suivantes : M1 = jext (Ri +e \u2013r) si (Ri < r < Rext) M2 = jext . e si (0 < r < Ri) avec \u2022 M1 : est l'aimantation ou le courant est non nul. \u2022 M2 : est l'aimantation ou le courant est nul. \u2022 Jext : densit\u00e9 de courant dans els bobines. \u2022 e : \u00e9paisseur de la bobine. Dans notre mod\u00e8le, l'\u00e9cran supraconducteur a \u00e9t\u00e9 d\u00e9fini comme un mat\u00e9riau ayant une perm\u00e9abilit\u00e9 relative tr\u00e8s faible, \u03bcr = 10-3. Les bobines supraconductrices ont \u00e9t\u00e9 dimensionn\u00e9es en tenant compte de la loi de Jc (B) du fil supraconducteur"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.13-1.png",
        "caption": "Fig. 35.13 Single and double row angular contact ball bearings.",
        "texts": [
            " Self-aligning ball bearings have two rows of balls and a common sphered raceway in the outer ring and this feature gives the bearing its self-aligning property which permits a minor angular displacement of the shaft relative to the housing. These bearings are particularly suitable for applications where misalignment can arise from errors in mounting or shaft deflection. A variety of designs are available with cylindrical and taper bores, with seals and adapter sleeves and extended inner rings. Angular contact ball bearings (Fig. 35.13) In angular contact ball bearings the line of action of the load, at the contacts between balls and raceways, forms an angle with the bearings axis. The inner and outer rings are offset to each other and the bearings are particularly suitable for carrying combined radial and axial loads. The single row bearing is of non-separable design, suitable for high speeds and carries an axial load in one direction only. A bearing is usually arranged so that it can be adjusted against a second bearing. A double row angular contact bearing has similar characteristics to two single bearings arranged back to back"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003465_ijamechs.2015.072817-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003465_ijamechs.2015.072817-Figure16-1.png",
        "caption": "Figure 16 BLDC core after winding wire coil with (a) winding speed at 100 rpm, (b) winding speed at 300 rpm, and (c) winding speed at 500 rpm; wire coils are numbered in clockwise direction (see online version for colours)",
        "texts": [
            " Moreover, the average of the wire tension error in the case of using ATS is below zero, in other word, the trend of wire is stretched due to the BLDC winding machine running at high speed winding generating high wire feeding acceleration. However, the wire tension error is reduced in the case of ATS compared to the case of PTS. The total time for winding three wire coils is about 70 seconds. The BLDC cores wound by the BLDC winding machine for three cases of winding speed and two tension systems of PTS and ATS are shown in Figure 16. The wire in bronze is fed by ATS while the wire in dark brown is fed by PTS. In all winding speeds, the wire coil fed by ATS has smaller size than the wire coil fed by PTS. Therefore, by equipping the active tension, it shows that the BLDC winding machine has ability to wind more wire turns to the wire coil. Besides, the wire coil in the case of the winding speed of 500 rpm is smallest in size owing to the trend of wire stretched. In summary, the comparison in wire coil volume approximated as well as volume reduction ratio between using PTS and ATS for feeding wire to the BLDC winding machine is listed in Table 3. In the BLDC core as shown in Figure 16, the 1st, 4th, and 7th wire coils are produced by using ATS, while the 2nd, 3rd, 5th, 6th, 8th, and 9th wire coils are produced by using PTS in winding process. Then, the reduction ratio of the volume of wire coil (RRV) is defined as follows: 0 0 100%v vRRV v \u2212 = \u00d7 (5) where v is the average volume of wire coil produced by using ATS, v0 is the average volume of wire coil produced by using PTS. The smaller the value of RRV is, the more the turn occurs on wire coil. Accordingly, the BLDC winding machine can manufacture the BLDC core smaller in ATS than in PTS"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003338_cdc.2016.7798877-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003338_cdc.2016.7798877-Figure3-1.png",
        "caption": "Fig. 3. Experimental high-precision stage and its model for simulation[15][16][17].",
        "texts": [
            " xreg d (t) is calculated using Eq. (27) by considering Eq. (19). xreg d (t) =  xreg 1d (t) xreg 2d (t) xreg 3d (t) ... xreg (n\u22121)d(t)  =  xreg 1d (t) d dt xreg 1d (t) d2 dt2 xreg 1d (t) ... dn\u22121 dtn\u22121 xreg 1d (t)  (27) F. State trajectory generation As abovementioned the state trajectory is obtained by xd(t) =  O (t < \u2212tpa) xreg d (t) (\u2212tpa \u2264 t \u2264 0) xst d (t) + xust d (t) (0 < t) . (28) V. Simulation results A. Simulation condition This section describes the simulations performed using the model illustrated in Fig. 3(b). This model assumes a highprecision stage shown in Fig. 3(a). Here, the continuous time domain transfer function is defined as Pc(s) = \u2212(s \u2212 140)(s + 100) s(s + 2000)(s + 2)(s2 + 20s + 40000) (29) assuming that the transfer function from the current reference of the x-axis actuator which generates force fx to the measured stage position x. \u03c4 is defined as the time constant of the unstable zero as \u03c4 = 1 140 \u2243 7.2 [ms]. (30) The discretized transfer function of Eq. (29) with zero-order hold is obtained as Ps[zs] = K(zs + 3.547)(zs \u2212 1.014)(zs \u2212 0.9900)(zs + 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000954_icems.2014.7013871-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000954_icems.2014.7013871-Figure2-1.png",
        "caption": "Fig. 2. Definition of coordinates",
        "texts": [
            " Accordingly, it is necessary that the suspension force is converted to the stator coordinate for displacement control. Equation (2) shows relationship between the suspension current and the suspension force in the stator coordinates. b a ddqqqd qqqddd mm mm i i iMiM iMiM PP PP F F 2 2 44,244,2 44,244,2 cossin sincos (2) where, F and F are suspension force in -axis of stator coordinates, i2a and i2b are suspension current on ab-axis of stator coordinates, is rotating angle of the rotor. The relationship between the rotor coordinates and the stator coordinates is shown in Fig. 2. In addition, (3) is obtained by solving the (2) in terms of the suspension current. F F iMiM iMiM PP PP i i ddqqqd qqqddd mm mm b a 1 44,244,2 44,244,2 2 2 cossin sincos (3) In the conventional control method, reference value of the suspension current is calculated by using (3). However, (3) shows that the suspension current in the stator coordinates becomes alternating current with high frequency in proportion to Pm times of rotating speed in the case of generating constant suspension force. For example, in the case of four-pole machine (Pm=2), 30000 min-1, frequency of the suspension current is 1 kHz. Because of increasing of frequency of the suspension current, suspension control in high speed rotation is difficult. Therefore, a method which is convert AC suspension current into DC by using a special rotating coordinates for the suspension windings has been proposed. Specifically, a\u02b9b\u02b9 rotating coordinate system which is rotating at Pm relative to the stator is defined as shown in Fig. 2. A conversion matrix from the stator coordinates to the a\u02b9b\u02b9 rotating coordinates is mm mm ab ba PP PP C cossin sincos'' . (4) In the (4), unlike in the case of normal dq transformation, the number of pole pair of motor winding Pm is used to the suspension windings with the number of pole pair Ps. By using (4) to (3), a relationship between the suspension current and the suspension force in a\u02b9b\u02b9 rotating coordinates is obtained as F F iMiM iMiM i i ddqqqd qqqddd b a 1 44,244,2 44,244,2 2 2 . (5) In the (5), terms associated with rotating angle of the rotor is canceled"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002948_gt2016-57458-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002948_gt2016-57458-Figure2-1.png",
        "caption": "Figure 2 Brazed design of the GT26 (Rating 2011 onwards) DFP design consists of 19 parts",
        "texts": [
            " CAD [-] Computer Aided Design CT [-] Computer Tomography DFP [-] Damper Front Panel FAA [-] Federal Aviation Administration FPI [-] Fluorescent penetrant inspection KIC [-] Stress intensity for plane strain deformation LCF [-] Low cycle fatigue Mred [kg (K)^1/2/s/bar] Reduced mass flow NDT [-] Non-destructive testing PT [-] Penetrant Test R [mm] Radius SLM [-] Selective Laser Melting TRL [-] Technology Readiness Level As stated in the previous chapter, the GT26 (Rating 2011 onwards) SEV Damping Front Panel (DFP) was chosen for validation of the SLM manufacturing technique in engine conditions based on the high degree of complexity of the nearwall-cooled channels and the high number of total subparts. The baseline product consists of 19 parts, where 3 main plates were brazed together with 16 brazing foils as shown in Figure 2. 2 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89513/ on 04/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use The idea of this study was to replace the complex baseline part by one single SLM part. The material used was Haynes 230. At the beginning of this study, small scope design modifications were performed for a quarter-section as well as for a full size trail part. The modified quarter-section and full size part were produced to verify not only SLM manufacturing capability, but also to check the design, regarding the range of necessary modifications"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000908_amr.1028.90-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000908_amr.1028.90-Figure1-1.png",
        "caption": "Fig. 1 Principle of Coaxial Powder Feeding Laser Cladding",
        "texts": [
            " The experiments of the laser cladding were performed using a 2 kW semiconductor laser. The FH-PFD auto-feeding powder equipment was used as the powder feeder which can provide the powder feeding voltage ranged from 6.5 V to 24 V. The laser beam was adjusted to a circular spot with size of 2mm. The powder nozzle is 1.5mm in diameter. The angle between the powder nozzle and the substrate was set about 45 degrees and the distance was set at 20 mm. The coaxial powder feeding was used for the experiments. Figure 1 [1] shows the principle of coaxial powder feeding laser cladding. Single track single layer experiments and multi-tracks single layer experiments were performed to study the effects of cladding processing parameters on cladding layers in this paper. The qualities of the cladding layer can be affected greatly by the laser cladding process. The processing parameters mainly include laser power, laser spot size, scanning speed, powder feed rate, overlap ratio and so on. In this paper, the laser spot size was adjusted to 2mm in diameter and kept as the constant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002089_gt2015-42624-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002089_gt2015-42624-Figure4-1.png",
        "caption": "FIGURE 4. CUT THROUGH MOSQUITO, (P) INDICATING PRESSURE MEASUREMENT LOCATIONS",
        "texts": [
            " The new experiment called \u201cMosquito\u201d, was designed to include exact flow path geometries of the modified turbine and to accommodate the original rotor and stator as well as the new seal components. The main difference between the full rotating rig and \u201cMosquito\u201d is the absence of the bearing chamber and a resulting stationary \u201cfrozen\u201d rotor. Additionally \u201cMosquito\u201d provides a high degree of flexibility and supplementary measurement points made much easier by the elimination of rotating elements. A cut through of a fully assembled \"Mosquito\" experiment is presented in Fig. 4. The new experiment still featured the original rotor and stator of the turbine only the rotor was now fixed in by the \u201cMosquito Core\u201d. The axial position of the rotor could be fine tuned using shims between the core and modified turbine shaft to vary axial clearances. Additional pressure measurement points were created between the rotor and stator through one of the bolts used to fix the stator as well as an additional measurement point in front of the stator inlet. These are marked on Fig. 4. The turbine outlet pressure was measured 300mm downstream of the rotor exit which corresponded to approximately 40 axial cord lengths of the rotor blade and therefore featured mixed out uniform flow. The secondary leakage path flow rate through the shaft seal was eliminated to better represent the set-up of 2D CFD investigations described above. During normal operational conditions the nozzles of the stage were always choked which means that the mass flow rate through the turbine was governed entirely by the pressure upstream of the stator. However it was felt essential to vary the pressure on the inlet and outlet of the rotor as well. In order to achieve this the stator was removed and the experiments presented in this paper were conducted on the set-up as shown in Fig. 4 excluding the nozzles. The \u201cMosquito\u201d experiment was subsequently integrated into the Durham Blow Down Facility as shown in Fig. 5. The Durham Blow Down Facility (DBDF) was designed to carry out a number of short duration high speed aerodynamic tests and has been used by a number of different projects including these experiments on fluidic sealing. The DBDF receiver is a 10m3 tank capable of storing air pressurised up to 30bar. For the present work the tank was operated at a maximum of 15bar. At 15bar the tank contained around 175kg of air which allowed the facility to run for up to 15 minutes assuming a minimum level of tank pressure of 11bar and an 4 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings",
            " Further to eliminate random fluctuations during a run, for each measurement point 2000 samples were taken and averaged with a sampling frequency of 1000Hz which placed the mass flow rate measurement uncertainty for each point well under 1%. The exception were the points with very low mass flow rate <0.001kg/s where the uncertainty suddenly increases up to 20%. This sudden increase is due to the mass flow calculations as described in BS EN ISO 5167-1/2. Simultaneously with the mass flow rate measurement the pressures on the experiment as indicated in Fig. 4 were mea- sured using a 16 channel multiple transducer ScaniValve module (Model DSA3217). The ScaniValve is accurate to 0.05% full scale for a range 0bar to 6.8bar. Here to minimise the random error 800 samples were taken and averaged at a sampling frequency of 400Hz. The ScaniValve module delivers values in engineering units. For the mass flow rate measurement each voltage supplied by the pressure transducer was logged using National Instruments (USB6218) logging card and processed using \u201cDurham Software for Windtunnels\u201d where calibration values were applied and pressures calculated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003465_ijamechs.2015.072817-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003465_ijamechs.2015.072817-Figure2-1.png",
        "caption": "Figure 2 Winding system (see online version for colours)",
        "texts": [
            " Copper/aluminium wire from a wire spool travels to a wire cleaner through a roll to roll system. Then, the wire is cleaned by two felt pads as well as anointed oil for reducing the friction force between the wire and machine components during wire transferring. To generate an essential wire tension force for winding process, a tension system acts on the wire before feeding to a winding part. The structure of the winding part includes a pneumatic system for clamping BLDC core and cutting the wire as well as a winding spindle in Figure 2 involving two cam mechanisms which is driven by three AC servo motors. Since a BLDC core has nine coil teeth in Figure 2 in order to reduce the time required for winding, three coil teeth are wound by three wires at the same time. Hence, a winding spindle employs three nozzles for driving three wires in winding process. The procedure of winding is described as shown Figure 2. The BLDC core is fixed between clamper and core bearer in Figure 1. Winding spindle equipped with three nozzles moves to up/down direction and swings to right/left direction due to the pilot of two cam-mechanisms in Figure 1. By this motion of the winding spindle, the end point of the nozzle with the wire wraps around the coil tooth formed into a wire turn. To construct a layer of wire turn, the nozzle is driven to move along the coil tooth side. This winding procedure is repeated until winding nine wire coils fully with four layers of the wire turn for each wire coil"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002813_b978-0-08-100072-4.00003-4-Figure3.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002813_b978-0-08-100072-4.00003-4-Figure3.4-1.png",
        "caption": "Figure 3.4 Implantable, self-contained titanium-housed battery, transmitter and sensor combination with two surface pads of polyester velour patches for fixation in tissue: (A) electronic sub-units, (B) telemetry transmission portal, (C) battery, (D) sensor location. With permission from Gough DA, Kumosa LS, Routh TL, Lin JT, Lucisano JY. Function of an implanted tissue glucose sensor for more than 1 year in animals. Science Translational Medicine 2010;2(42).",
        "texts": [
            " A range of other porous and homogeneous membrane formulations have been investigated [48], but it is unclear at this stage if an optimally advantageous material will emerge. Current semi-implantable commercial devices are designed to operate for limited periods, typically 3\u201310 days. For long-term monitoring, full surgical implantation would be necessary and the potential denaturation of enzyme by sustained exposure to the H2O2 product of an oxidase reaction, for example, would need to be considered. As a development from their earlier work on oxygen sensor-based intravascular electrodes, Gough\u2019s Group have developed a pillbox-type sensor (Fig. 3.4) using an oxygen cathode sensor [86]. The system has the advantage of monitoring the oxidase reaction using O2 changes and so can incorporate catalase in the enzyme layer to eliminate H2O2. The critical oxygen-sensing membrane is gas-permeable polydimethylsiloxane (PDMS) over an electrolyte film. Over this is arrayed a series of wells containing immobilised glucose oxidase; the geometry of the PDMS wells is such as to present a higher surface area for oxygen transport, as opposed to reduced diffusive access for glucose, so avoiding oxygen limitation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure9-1.png",
        "caption": "Figure 9 Schematic diagram that shows the calculations of deviations.",
        "texts": [
            " Figure 8 shows the loci of the cutter edge calculated for three rotations of the workpiece, in which the Lagrange interpolation was applied. Deviations are the distance between the true involute flank and a new flank in the normal direction of true involute flank. Therefore, unit normal vector N of the true involute flank Rinv is calculated at first. Next, calculate multiple cutting edge loci obtained from multiple rotations (n times) of the workpiece. By solving equation Rinv + tiN = Rlocus,i, deviation between the ith cutting edge and the true involute flank ti is calculated (Fig. 9). Iterate this calculation for nth cutting locus, and the minimum on them is the deviation at that point. By conducting this iterate calculations over the tooth height, one can obtain profile deviations. Similarly, by calculating over the face width, lead deviations are determined. Analysis under two conditions, that is, a) a cutter without pitch deviation and without run-out, and b) a cutter with a pitch deviation of 5 m and a run-out of 12 m (which is the same as the case of the experiment explained later) were carried out"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.6-1.png",
        "caption": "FIGURE 6.6",
        "texts": [
            " The trajectory (path) of the vehicle can also be recorded. With simulation this is straightforward but in the past has been difficult to measure on the test track; testers resorting to measuring a trail of dye left by the vehicle on the test track surface. Modern instrumentation has improved on this somewhat. Another measure often determined during test or simulation is the body slip angle, b. This being the angle of the vehicle velocity vector measured from a longitudinal axis through the vehicle as shown in Figure 6.6. The components of velocity Body slip angle. of the vehicle mass centre, Vx and Vy, measured in vehicle body reference frame can be used to readily determine this. In Chapter 4 the modelling and analysis of the suspension system was considered in isolation. In this section the representation of the suspension as a component of the full vehicle system model will be considered. As stated the use of powerful MBS analysis programs often results in modelling the suspension systems as installed on the actual vehicle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003151_s11837-016-2190-9-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003151_s11837-016-2190-9-Figure2-1.png",
        "caption": "Fig. 2. Test specimen according to DIN 50125, type B5 9 25. The dark marks show the measurement positions at h = 0 and h = 45 .",
        "texts": [
            " The initial load was set to 1 kN, and an extensometer was used for the microstrain measurement within the elastic range. The three-dimensional surface analyses were performed with a confocal microscope with a 20 9 0.6 objective and a resolution of 3.02 nm vertical and 1.57 lm lateral. Each measurement area was set to a size of 0.8 9 2.14 mm2 (Fig. 1) In total, nine measurement areas were analyzed for each test specimen, separated into the polar angles h = 0 , h = 45 , and h = 135 . The areas were located within the reduced section of the specimens (Fig. 2) For the evaluation of the three-dimensional measurements, in a first step, a form removal of polynomial degree two was performed. Next, the noise was removed with an S-filter with a nesting index of 0.008 nm according to the lateral resolution of 1.57 lm and DIN EN ISO 25178-3. Additionally, nonmeasured points were interpolated. The preliminary result is a form removed surface texture that consists of roughness and waviness shares. For a detailed analysis, those shares were separated into roughness (S\u2013L-surface) and waviness (S\u2013F-surface) textures with a nesting index of 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002901_stab.2016.7541212-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002901_stab.2016.7541212-Figure1-1.png",
        "caption": "Fig. 1. An axial-symmetrical finned body.",
        "texts": [
            " These intervals are compared for bodies with different aerodynamic properties of blades taking into account displacement of a center of mass. Conditions of stability of stationary motion of mechanical systems described by complex differential equations of the third order [1] are used for this purpose. Character types of descent of the body are shown taking into account displacement of a center of mass. A dynamically symmetrical finned body performs a free descent into an unperturbed atmosphere. The body is under the action of gravity and aerodynamic forces. The body has four similar blades symmetrically located on it (Fig. 1). Assume that aerodynamic action is concentrated on blades, and distributed system of aerodynamic forces acting on each blade is equivalent to a resulting force applied at a center \ud835\udc42\ud835\udc56 of the blade (center of pressure). The quasi-static model of an aerodynamic action is used [2]. This model allows obtaining of satisfactory description of nonstationary motions in a certain conditions [3]. Due to the symmetry of the problem, there exists a trivial stationary motion (autorotation) for which the axis of symmetry of the body is vertical, the center of mass of the body moves along the axis of symmetry with a constant speed V, angular speed of autorotation about the axis of symmetry is a constant value \u03a9"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001892_gt2015-42580-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001892_gt2015-42580-Figure2-1.png",
        "caption": "FIGURE 2. Main test arrangements (A: blow-down; B: axial behavior of the bristle pack; C: stiffness)",
        "texts": [
            " During the measurements the seal carrier is fixed. The focus at the test facility is directed to brush seal quality attributes, like the bristle pack stiffness, the blow-down or the axial behavior of the bristle pack. Additionally, the leakage flow can be metered for all test arrangements. Furthermore, the test facility provides optical access from the downstream side for all test arrangements to the bristle pack which offers the possibility to record and analyze sealing behavior. An overview about the complete test facility is given in [6]. Figure 2 depicts the three main settings of the test facility to analyze the brush seal behavior. Set-up \u2019A\u2019 is used to determine the blow-down capability of the test seal. Therefore, radial clearance (cg) between the last bristle row and the shaft disc is photographed and analyzed for different pressure drops. 2 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 04/26/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use During the test, the leakage flow through the seal is measured simultaneously",
            " Set-up \u2019B\u2019 allows analyzing the axial packing and the axial behavior of the bristle pack. Therefore, an optical access including a mirror is assembled to the shaft disc, which enables photographs or motion pictures to be taken from below at different pressures. Measuring the leakage flow, the set-up \u2019A\u2019 and \u2019B\u2019 are used to check that the data matches for the test seal. By measurement of the leakage flow the identified behavior and performance of the tested seal in set-up \u2019A\u2019 and \u2019B\u2019 can be ensured. A further set-up which is shown in Fig. 2 \u2019C\u2019 is used to test the radial stiffness of a bristle pack segment. Thereby, a motor controlled movable test shoe is installed at the shaft disc including a force and displacement sensor. For testing, the test shoe is moved 0.5 mm radial through the bristle pack at certain pressure drops. As discussed in [13] stiffness testing by using a test shoe includes non-linear end effects which influences the absolute stiffness determination for a brush seal. Nevertheless, a comparison of different test arrangements is possible since the results from different arrangements contain similar end effect errors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002450_s11029-016-9566-3-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002450_s11029-016-9566-3-Figure2-1.png",
        "caption": "Fig. 2. Schematic sketch of the experimental unit. Explanations in the text.",
        "texts": [
            "4) will be used in Sect. 4 to identify the shear damping properties of the CFRP in the layer plane. For determination of the dynamic elastic properties and the amplitude dependence of damping properties of materials at low frequencies, test specimens of significant length have to be used. To exclude the static component of deflection, it is reasonable to perform dynamic tests of such specimens in their vertical position. For this purpose, an earlier developed experimental setup [2], whose schematic is shown in Fig. 2, was updated. The unit consists of a base 1 and a load-bearing rack 2 rigidly connected together. On the rack, a cantilever 3 with a grip 4 on its end is fixed stationary. The test specimen 5 is fixed by spaced rigid plates connected to the cantilever by bolt joints so that to suppress rotation of the test specimen in the cross section of restraint. On the rack, a mobile platform 6 is mounted for fastening a laser displacement gage 7, whose position along the rack can be varied to measure the vibration amplitude of the end of test specimen upon changing its arm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000212_978-981-15-5712-5-Figure17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000212_978-981-15-5712-5-Figure17-1.png",
        "caption": "Fig. 17 A single-axis cell stretcher in unstretched and stretched configurations (adapted from [27]) using the spring shown in Fig. 16",
        "texts": [
            " Generating a straight-line motion with the help of only hinge joints in a mechanism was a design challenge and Watt considered this mechanism to be his best work. Slowly the subject of mechanism theory started achieving maturity as the steam engines started taking modern shapes with the introduction of high-pressure steam engines by Richard Trevithick in 1801. For controlling the speed, a governor mechanism was also introduced by James Watt, and the centrifugal governor designed by him is still called \u2018Watt Governor\u2019. Figure 16 shows Watt\u2019s design of the centrifugal governor he employed. Figure 17 shows a double-acting rotary engine constructed in 1788. It shows the introduction of the straight-line mechanism and the governor. Figure 18 shows the drawing of a Watt\u2019s double-acting rotary engine. Since the piston moved in a straight line, in this case, its connecting point driving the rocking beam needed rectilinear motion. The use of the straight-line mechanism and a Watt governor is clearly visible. The sun and planet mechanism to convert oscillatory motion to unidirectional continuous rotarymotion is also conspicuous in the drawing",
            " This is because what we see here is an array of springs in series wherein each spring is a pair of two fixed-guided beams forming a rectangular box. It is indeed economical use of material and manufacture. Such a stack of springs has many uses. Figure 16b shows two such springs (white and gray colors interchanged from Fig. 16a with white representing beams here) with a central ring. If the central ring is moved to one side, one spring expands and the other contracts. This structure was used as a cell stretcher [27]. As illustrated in Fig. 17a, b, when biological cells Fig. 16 An example of economy material and manufacture: a a very flexible spring is realized by cutting out a few slits (white lines); b a stack of two such springs on either side of the central ring The Art and Signs of a Few Good Mechanical Designs in MEMS 43 are seeded on a patterned polymer layer, SU-8, in particular, cells are stretched by applying a force to stretch the spring. In fact, the stiffness of the spring is so low that cells themselves would stretch the spring, paving the way to measure forces applied by the cells [27]",
            " Though the generalized computerization allows arbitrary choices of pc and pd for any given case study, such arbitrary choices show that there is a threshold interpolation order at which numerical instability becomes overbearing. Stability curves are generated for pc = pd = p > 10 and compared against the reference in Figs. 19 and 20. It is seen that pc = pd = p = 9 is the threshold interpolation order beyond which the FDM fails for the studied system. The failure starts from the high speed domain and extends deeper into the low speed domain for higher orders p, and the magnitude of the failure is higher for higher orders p at every spindle speed. Fig. 17 The identified stability curves at two-thirds of tool pass of the bidirectional model for the orders p = pc = pd 0.5 1 1.5 2 2.5 spindle speed [rpm] 104 0 5 10 15 20 25 30 de pt h of c ut [m m ] ref p=1 p=2 p=3 p=4 p=5 p=6 p=7 p=8 p=9 p=10 CTRS Fig. 18 The time-domain simulation of nodal responses at [15000 rpm, 22.5 mm] indicated by star in Fig. 17 0 0.01 0.02 time [s] -400 -300 -200 -100 0 100 200 300 re ge ne ra tiv e di sp la ce m en t [ m ] 0 0.01 0.02 time [s] -1 -0.5 0 0.5 1 re ge ne ra tiv e ve lo ci ty [m /s ] Fig. 19 Stability curves for p = pc = pd compared against the reference at the start of tool pass 0.5 1 1.5 2 2.5 spindle speed [rpm] 104 0 5 10 15 20 25 30 de pt h of c ut [m m ] ref p=11 p=12 p=13 p=14 p=15 Fig. 20 Stability curves for p = pc = pd compared against the reference at two-thirds of tool pass 0.5 1 1.5 2 2.5 spindle speed [rpm] 104 0 5 10 15 20 25 30 de pt h of c ut [m m ] ref p=11 p=12 p=13 p=14 p=15 A method which combines modal truncation and tensor-based general order FDM was developed to suppress all the case-by-case symbolic analyses associated with stability analysis of elastic thin-walledworkpiece"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003119_icelmach.2016.7732831-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003119_icelmach.2016.7732831-Figure3-1.png",
        "caption": "Figure 3: Graphical presentation of the used bearing parameters",
        "texts": [
            " Simplifications of the system where carried out. The outer race and the inner race are represented by two solid profiles which are able to translate with respect to each other. Between those profiles, only one bearing ball is drawn which will move due to an implied translation and rotation. Several pits are added in order to simulate the other bearing balls rolling over the same pit. The specified distance between the pits is in fact the distance between the bearing ball contact points with the outer race (Figure 3). This distance dout can be expressed as: dout = \u03c0 m (pd+ bd \u00b7 cos\u03b2) (1) With: dout distance between pits in simulation [m]; m number of bearing balls (8); pd pitch diameter (74mm); bd ball diameter (17.5mm); \u03b2 contact angle (0\u25e6). The bearing ball has the dimensions and the mass of the original IM\u2019s replaced bearing used in the test-rig, superposed with the mass corresponding to the gravitational force which was working on that IM\u2019s bearing at DE (mrot,DE). This implies that all inertia properties of the ball in every direction remain the same except the inertia in the vertical direction (perpendicular on the direction of movement), which is now related to mrot,DE",
            " A bearing inner race fault seems in principle very similar to an outer race fault. Nevertheless, the fault propagation is to complex to simulate with the mechanical FEM software. With the use of previously derived system parameters of the SDOF interaction between the stator and the rotor, the vertical rotor movements can be obtained analytically. The velocity of the inner ring with respect to the ball velocity is: vball,in = \u03c0 (pd\u2212 bd \u00b7 cos\u03b2) (frot \u2212 fFTF) (9) The relative distance, dinn, between the different contact points of the balls with the inner race (Figure 3) can be expressed as: dinn = \u03c0 m (pd\u2212 bd \u00b7 cos\u03b2) (10) With the use of vball,inn and dinn, the pulse train frequency, vball,inn dinn , is calculated, presented in Figure 6. The magnitude of the impulse function (450N) is based on the reversely calculated impulse function of the previously modeled outer race bearing fault. The impulse function of that outer race fault is mathematically derived by performing the deconvolution of the response function (Figure 4) and the impulse response function (analytically obtained with the system parameters mrot,DE, krad,DE and crad,DE)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002944_icma.2016.7558946-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002944_icma.2016.7558946-Figure2-1.png",
        "caption": "Fig. 2. Model construction. Geometry model closely envelope the topology network, and the triangular patches in the geometry model is follow the nodes\u2019 motion.",
        "texts": [
            "90mm 435 TABLE II PRARAMETERS OF GEOMETRY MODEL The vascular geometry model is built with the basic element of triangular patch. The parameters of the geometry model show in table II. The less the number of triangular patches, the faster calculation speed, but the worse resolution of the model. So, the calculation speed and models\u2019 resolution must be take into account simultaneously. The strategy of geometry model closely envelope the topology network is used to construct the vascular model, as shown in Fig. 2. In this figure, each point in topology network represents a mass element, and each line represents a damping spring connecting two mass elements. The deformation of the vascular model in the gravity field is shown in Fig. 3. In Fig. 3a, the vascular model\u2019s cross section deformed in an elliptical shape under the effects of gravity, and the model has an obvious refractive effect, as shown in Fig. 3b. The biomechanics properties of biological vascular substitutes(BVS) in Francois Auger\u2019s laboratory were tested by Georgia Institute of Technology [8]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003232_detc2016-59194-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003232_detc2016-59194-Figure4-1.png",
        "caption": "Fig. 4 A multi-mode 7R mechanism.",
        "texts": [
            "org/about-asme/terms-of-use It is noted that the explicit equations Eqs. (31) and (41), (37) and (40) as well as (38) and (39) are, respectively, in the same form. This confirms the symmetric characteristics of the line symmetric 7R mechanism. 7R SPATIAL MECHANISM In this section, the loop equation Eq. (21) in section 3 will be used for the reconfiguration analysis of multi-mode mechanisms. The reconfiguration analysis of a multi-mode 7R mechanism will be discussed in detail. A multi-mode 7R mechanism shown in Fig.4. This mechanism was obtained using the construction method [14] by integrating an orthogonal Bricard mechanism 1-2-3-4-5-6 with a planar 4R mechanism 1-3-5-7 with three joints (1, 3 and 5) in common. The link parameters of the mechanism are: di = 0, i = 1, 2, . . . 7 Li = 2, i = 1, 2, . . . 5 L6 = L7 = 1 \u03b11 = \u03b13 = \u03b15 = \u03c0/2 \u03b12 = \u03b14 = \u03b16 = \u2212\u03c0/2 \u03b17 = 0 Using Eq. (21), we obtain the following set of kinematic loop equations  e1(u1, u2, \u00b7 \u00b7 \u00b7u7) = 0 e2(u1, u2, \u00b7 \u00b7 \u00b7u7) = 0 e3(u1, u2, \u00b7 \u00b7 \u00b7u7) = 0 g1(u1, u2, \u00b7 \u00b7 \u00b7u7) = 0 g2(u1, u2, \u00b7 \u00b7 \u00b7u7) = 0 g3(u1, u2, \u00b7 \u00b7 \u00b7u7) = 0 (42) where the detailed expression of these equations are omitted for space reason"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure5.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure5.8-1.png",
        "caption": "Fig. 5.8 Influence of the coordinate system\u2019s origin",
        "texts": [
            " Strong formulation LTCLu0 + b = 0 Inner product\u222b V WT(x) ( LTCLu + b ) dV = 0 Weak formulation\u222b V (LW)T C (Lu) dV = \u222b A WT t dA + \u222b V WTb dV Principal finite element equation (quad 4) \u222b V ( LNT )T C ( LNT ) dV \ufe38 \ufe37\ufe37 \ufe38 K e \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 u1x u1y . . . u4x u4y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 F1x F1y . . . F4x F4y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 + \u222b V N b dV 5.3.3 Solved Plane Elasticity Problems 5.1. Example: Influence of the coordinate system\u2019s origin on the geometrical derivatives Given is a square two-dimensional element as shown in Fig. 5.8. Calculate the geometrical derivatives of the natural coordinates (\u03be, \u03b7) with respect to the physical 5.3 Finite Element Solution 259 coordinates (x, y) and consider the different locations of the elemental coordinate system as shown in Fig. 5.8a\u2013c. Comment: the parametric \u03be\u03b7-space as shown in Fig. 5.8d would not allow this flexibility since \u22121 \u2264 \u03be \u2264 1 and \u22121 \u2264 \u03b7 \u2264 1 must hold. This problem relates to steps \u2776 to \u2779 as given on p. 256 and 257. 5.1. Solution The coordinates of the four corner nodes in the different xy-systems are collected in Table5.6 Thenext step is to calculate the partial derivatives of the physical (x, y) coordinates with respect to the parametric (\u03be, \u03b7) one, see Eqs. (5.67)\u2013(5.70). For case (a), this evaluation gives: \u2202x \u2202\u03be = 1 4 ( (\u22121 + \u03b7)(\u2212a) + (1 \u2212 \u03b7)(a) + (1 + \u03b7)(a) + (\u22121 \u2212 \u03b7)(\u2212a) ) , \u2202y \u2202\u03be = 1 4 ( (\u22121 + \u03b7)(\u2212a) + (1 \u2212 \u03b7)(\u2212a) + (1 + \u03b7)(a) + (\u22121 \u2212 \u03b7)(a) ) , \u2202x \u2202\u03b7 = 1 4 ( (\u22121 + \u03be)(\u2212a) + (\u22121 \u2212 \u03be)(a) + (1 + \u03be)(a) + (1 \u2212 \u03be)(\u2212a) ) , \u2202y \u2202\u03b7 = 1 4 ( (\u22121 + \u03be)(\u2212a) + (\u22121 \u2212 \u03be)(\u2212a) + (1 + \u03be)(a) + (1 \u2212 \u03be)(a) ) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001030_haptics.2014.6775481-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001030_haptics.2014.6775481-Figure3-1.png",
        "caption": "Figure 3: Target force responses: hard and soft rubber sheet",
        "texts": [],
        "surrounding_texts": [
            "In this paper, we describe the haptic rendering method of an insertion task with a needle. We employed a new haptic rendering method in which the virtual force generated from a haptic device is superimposed on the force response of a base object. A rubber sheet was used as the base object. An actuator stretched the rubber sheet in order to change the elasticity of the sheet and the transition displacement of the force response. The experimental results showed that the proposed haptic augmentation can display the force response during the needle insertion task into the target rubber sheet.\nIndex Terms: H.5.2 [INFORMATION INTERFACES AND PRESENTATION]: User Interfaces\u2014Haptic I/O;\nHaptic rendering technologies have long been proposed as a function that interfaces should have for the intuitive transmission of information from a system to a user. There are many studies in this field such as tele-operation[?] and rehabilitation[?] reflecting a strong need for haptic rendering. Methods for calculating force responses in the deformation of an object have also been studied [?, ?]. Some studies have reported the haptic information is very useful in training surgical skills [?, ?]. Mechanical properties of a variety of human organs and tissue have been measured [?], and these properties were used to calculate the force response when manipulating virtual human tissue [?]. In these studies, highprecision modeling of the force response requires significant calculation time and high-performance devices to fully render the haptic senses. With poor performance of the model or the device, users feel a sense different from the one they should feel.\nOur research group have proposed the use of force rendering by augmentation to realize a realistic force response even using a lowperformance haptic device [?]. The key idea is to use an object that has material properties similar to those of a target object, and to overlap the force produced by the haptic device on the force of the base object as shown in Fig. ??. The haptic senses produced by the proposed Haptic Enhanced Reality (HER) method are improved because the base object covers a haptic response that we cannot perfectly model or that the device cannot completely produce.\nIn this paper, we explore the feasiblity of the HER method to a needle insertion task. The needle insertion task is a basic manipulation for the manual skill training in the fields of nursing and medicine. We first explain the outline of the HER method and provide a general discussion of the characteristics that the base object should have. Next, we explain the method to change the elastic property and the tear resistance of a rubber sheet by stretching. The force response of the rubber sheet can be controlled by changing the\n\u2217e-mail: kurita@bsys.hiroshima-u.ac.jp\nintensity of stretching. We also describes the experimental results to confirm the efficacy of the proposed method.\nUsing a simple formulation, we first explain the characteristics of the haptic enhanced reality (HER) method. The force response of a target object Fdesired is expressed by the model force Mtarget and the model error Etarget :\nFdesired = Mtarget +Etarget . (1)\nThe model error should be small when model accuracy is high, although the formula for obtaining Mtarget would become complicated in this case. When Fdesired is produced by a haptic device alone, the force displayed to a user Fvr should be equal to the output force of device Fdevice:\nFvr = Fdevice. (2)\nWhen Fdevice is determined by the model force, we have\nFdevice = Mtarget . (3)\nThe difference between the target force Fdesired and the displayed force Fvr is given by\nFdesired \u2212Fvr = Etarget . (4)\nThis indicates that reducing the model error is necessary to improve the rendering accuracy.\nIEEE Haptics Symposium 2014 23-26 February, Houston, Tx, USA 978-1-4799-3131-6/14/$31.00 \u00a92014 IEEE",
            "Next, assume that the force displayed to the user Far is determined from the total force of the device output Fdevice and the reaction force of the base object Fbase:\nFar = Fdevice +Fbase. (5)\nThe reaction force of the base object is modeled as well:\nFbase = Mbase +Ebase. (6)\nThe difference between the target force Fdesired and the displayed force Far is then given by\nFdesired \u2212Far = Ftarget \u2212 (Fdevice +Mbase +Ebase). (7)\nWith Eqs.(1) and (7), we have\nFdesired \u2212Far = Mtarget \u2212Mbase +(Etarget \u2212Ebase)\u2212Fdevice. (8)\nIf the device produces the force that corresponds to the difference between the model force of the target and that of the base object, we get\nFdevice = Mtarget \u2212Mbase (9)\nError between the displayed force and the target force is therefore finally given by the following expression:\nFdesired \u2212Far = Etarget \u2212Ebase. (10)\nThis equation indicates that the rendering accuracy can be improved if model error of the base object cancels that of the target. In other words, as the model errors of the base object Ebase and the target object Etarget are closer, the desired force Fdesired and the displayed force Far get closer. Taking this into consideration, we propose the control method of the force response of the base object by actuating it.\nThe aim of this study is rendering the haptic force during an insertion task by using the aforementioned HER method. Fig. ?? shows a typical force response when a needle is penetrated into a rubber sheet. The force response during the needle insertion can be divided into three phases:\n1. the pushing force from the needle and the reaction force of the rubber sheet are balanced.\n2. the needle is tearing the rubber sheet.\n3. the needle passes through the rubber sheet.\nIn the phase (1), the reaction force of the rubber sheet acts like a spring. It monotonically increases according to the displacement. In the phase (2), once the force from the needle exceeds the limit of the tear resistance of the sheet, the needle is going to tear the sheet off. In the phase (3), when the needle passed through the sheet, the reaction force is going to keep a static value or be gradually increased depending on the depth of the insertion. The transitions from (1) to (3) phases play an important role to create the sense of the insertion. Based on these observations, we modeled the force response of the needle insertion by using the three force models.\nThe force response during the needle insertion changes depending on the hardness of the rubber sheet. Fig. ?? shows the force response during the needle insertion for soft and hard rubber sheets. Generally, a harder sheet shows larger response force during the pushing phase, and it reaches the limit of the tear resistance earlier. In contrast, a softer sheet shows smaller force, and large deformation is required to reach the limit.\nIn order to render the force during the insertion by using a base object, the displacement of the transitions between the force phases, which are shown in Fig. ??, is needed to be controlled. In this study, a soft rubber sheet with Asker C hardness of 40 and the thickness of 0.5 [mm] was used as the base object. In addition, the base object was stretched by an actuator to change the elastic property.\nFig. ?? shows the pictures when the base object is stretched 15 [mm] to both directions, totally 30[mm]. The elastic property of a rubber sheet can be changed by stretching the sheet; the transition displacement of the force phases can be controlled to match with the desired transition displacement. Fig. ?? shows the force response when the rubber sheet is streached 10, 20, and 30 [mm] from the default position (0 [mm]). The reaction force was statically measured for every 0.5 [mm] when the needle was vertically pressed to the rubber sheet. The reaction force increases until the force reaches the limit. The limit of the tear resistance does not change depending on the stretch; tearing occurs when around 0.25",
            "[N] force is applied. However, the pushing displacement required to reach the limit is different: 9 [mm] displacement for the default position, 7 [mm] displacement for 10 [mm] stretch, 5 [mm] displacement for 20 [mm] stretch, and 4 [mm] displacement for 30 [mm] stretch are required. This indicates that the transition displacement between the force phases can be controlled by changing the stretch intensity of the sheet.\nAs described above, it is possible to control the transition displacement between the force phases by stretching; however, the reaction force also changes depending on the stretching intensity. In order to render the desired force, a haptic display device generates an additional force and enhances the reaction force of the base object.\nBased on Eq. ??, the force that the haptic device generates Fdevice was determined by the difference between the model force of the target object Mtarget and the model force of the base object Mbase. In this study, Mtarget and Mbase were characterized as following:\nMbase = abase 1 x2 +abase 2 x (11)\nMtarget = atarget 1 x2 +atarget 2 x (12)\nwhere x is the pushing displacement, and a1,a2 are the coefficients. Fig. ?? shows the model force calculated by Eq. ??, and the measured static force response of the soft rubber sheet. The parameters used in the calculation are shown in Table ??.\nThe force that the haptic device generates Fdevice can also be determined by Eq. ??:\nFdevice = (atarget 1 \u2212abase 1 )x2 +(atarget 2 \u2212abase 2 )x. (13)\nFig. ?? shows the overview of the force display system. The system was composed of the haptic display device (Sensable: Phantom Omni), a rubber sheet as the base object (Asker C hardness: 40, thickness: 0.5 [mm]), a one-axis linear actuator, and a needle with the diameter of 0.5 [mm]. The target object was the hard rubber sheet with the Asker C hardness of 60 and the thickness of 0.5 [mm]. The force response during the needle insertion was statically measured by a one-axis force transducer and a one-axis displacement sensor every 0.5 [mm] displacement. The solid line in Fig. ?? shows the force response of the target rubber sheet during the needle insertion. In order to display this force response by the HER"
        ]
    },
    {
        "image_filename": "designv11_64_0002529_s106345411602014x-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002529_s106345411602014x-Figure5-1.png",
        "caption": "Fig. 5. Oscillations of the first pendulum.",
        "texts": [],
        "surrounding_texts": [
            "The results of numerical computations obtained with using the old and new approaches for the motion of a trolley with pendulums were compared to each other. Controlling force F and rotation angles of pendulums were calculated as functions of time. The force was expressed in Mg fractions where M is the mass of the whole system and rotation angles \u03c61 and \u03c62 were expressed in degrees. As was noted above, these functions, defined as functions of dimensionless time \u03c4 = t/ , depend significantly on parameter K, which is equal to the ratio of dimensionless time of movement T to the period T1 of the first form of oscillations of the mechanical system. The case when the value of K is between one and two [11] is of particular interest. Taking this into account, during the calculations it was considered that K = 1.54. The sought-for functions are proportional to the trolley movement S in time , so this movement expressed in fractions of length of the longest pendulum was chosen so that the pendulums rotation angles would not exceed ten degrees. In Figs. 2\u20137, the solid lines correspond to results obtained with the new approach, and the dashed ones correspond to the results of the old method. Figures 2 and 3 correspond to the calculations when the trolley hosts one pendulum; the system parameters were set as follows: T T VESTNIK ST. PETERSBURG UNIVERSITY. MATHEMATICS Vol. 49 No. 2 2016 A NEW APPROACH TO FINDING 189 190 VESTNIK ST. PETERSBURG UNIVERSITY. MATHEMATICS Vol. 49 No. 2 2016 ZEGZHDA et al. and for Fig. 7 The calculations predictably showed that the bigger the mass of the trolley compared to the pendulums, the closer the results of the first and the second approach. Therefore the trolley mass was selected to be half the mass of the first main pendulum. The length and the mass of the second pendulum in relation to the first one were selected to be 1/4 and 1/8, respectively. The trolley movement is equal to 0.2l1, where l1 is the length of the main pendulum. We see that application of the new approach results in reduction of the pendulums oscillations both in amplitude and in frequency. Comparison of Fig. 4 and 2 shows that for a new approach the addition of the second pendulum affected controlling force insignificantly, while for the usual approach it changed significantly. The following result is also curious. In the problem with two pendulums, let the pass of the second one be not 1/8, but rather 1/64 of the first one. A comparison of Figs. 7 and 4 shows that, for the new approach, such a substantial decrease in mass of the second pendulum barely affected the control force, while for the usual approach, it changed significantly."
        ]
    },
    {
        "image_filename": "designv11_64_0002773_9781782421955.1049-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002773_9781782421955.1049-Figure5-1.png",
        "caption": "Figure 5 \u2013 a) Support plate sub-assembly and FRF input/output points for one representative FRF, b) Bearing constraint detail.",
        "texts": [],
        "surrounding_texts": [
            "a deflection of a shaft half until the gear body. Another difference noticed for the shaft sub-assembly is a variation of the oscillation planes for the bending modes. While in the case of bare shaft all the bending mode pairs were oscillating only on two orthogonal planes, in the shaft sub-assembly each mode pair still oscillate on two orthogonal planes, however these planes have different orientations from pair to pair (Figure 4). The gear accelerometers carrier is suspected to be the cause of such effect, in fact this component has two preferential planes of symmetry which maximise and minimise bending stiffness. Namely, the test shaft sub-assembly without accelerometers carrier approximates well fully-cylindrical symmetry, while the accelerometers carrier introduces a 4-fold symmetry.",
            "and the displacement has been measured with a dial indicator. The first mode of the shaft sub-assembly dramatically changes both in frequency and shape. In unloaded conditions, the natural frequency is 695 Hz and the mode shape free-free like; in loaded conditions, the natural frequency drops to 481 Hz (-30%) and the mode shape becomes of hinged type. Similar observations can be anticipated for gears, if these can lose contact and move in the backlash. Following these observations, modal analysis has been performed in loaded conditions. Torque has a stiffening effect on the elastomeric couplings; stiffness rises from 1000 Nm/rad at 50 Nm to 5000 Nm/rad at 350 Nm. For the current discussion, the torque preload in the system has been fixed at 300 Nm, although modal analysis will be repeated at different torque levels. Impact testing showed limitations due to accessible surfaces and maximum frequency range that could be excited with sufficient energy, therefore modal analysis has been performed by shaker testing. The application of the force proved to be of critical importance and three configurations have been evaluated. The first location has been selected trying to excite the assembly as close as possible to the tooth mesh. A small eye bolt plate has been screwed to the gear body and a force has been applied normal to the plate surface in the direction of the Line of Action (LOA) (Figure 8a). The measured FRFs and the calculated mode shapes are not reliable, because they appear to be dependent on the stinger length and on resonances of the shaker fixture. The dependency happens because the force cell does not fully decouple the shaker fixture from the tested assembly: despite the efforts, the force cell could only be placed at a small distance from the gear body and the elasticity of this segment introduces stray resonances. Having the aim of exciting the system with a force correctly measured, the second input point location has been moved to the cylindrical surface of the gear locking element, with a vertical direction intersecting the shaft axis (Figure 8b). It is still possible to excite torsional modes of the system, thanks to the torsional-lateral coupling introduced by the meshing gears. With this configuration, the shaker fixture is correctly isolated by the force cell. However, only for the shaft excited indirectly (on the left in figure 8a-b), the obtained mode shapes show dominant motion along the LOA. This unwanted effect is attributed to anisotropy related mainly to the gear mesh, which can transmit forces orthogonal to the LOA only by friction. Therefore a force in the vertical direction can excite the shaft directly in both the LOA and the orthogonal directions, while friction is likely to filter the force in the orthogonal direction for the other shaft through the gear mesh. A secondary source of anisotropy is expected to stem from bearing stiffness, the\nClose to tooth mesh a); on gear clamping element in a vertical direction b)\nand aligned with the line of action c-d)."
        ]
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure8-1.png",
        "caption": "Figure 8. Three claws B.",
        "texts": [
            " In order to lock Shaft C in clockwise and counterclockwise directions, each drive unit has two velocity-based mechanical safety devices (that is, one velocity-based safety device for locking in the clockwise direction and another velocity-based safety device for locking in the counterclockwise direction). A. Velocity-based Safety Device In this section, we review the structure and mechanism of the velocity-based mechanical safety device briefly (see [13] for more information). Fig. 7 shows the structure of the safety device. Gear A, Plate B and Ratchet Wheel C are attached to Shaft C. Plate C has inner teeth. Each Claw B is attached to Plate B by Pin B and positioned as shown in Fig. 8. Guide Bar B is attached to each Claw B. Three Guide Bars B are respectively inserted into three Guide Holes A of Plate A. Plate A has ratchet teeth. Shaft C rotates Plate A via Torsion Spring. Three Claws B normally rotate together with Shaft C. One end of Linear Spring C is connected to Pin C1 attached to Plate C, and another end is connected to Pin C2 of Frame C. Each Claw C has Guide Bar C. Four Guide Bars C are respectively inserted into four Guide Holes C of Plate C. Gear D meshes with Gear A"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001493_s0021894414050162-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001493_s0021894414050162-Figure1-1.png",
        "caption": "Fig. 1. Setup with free oscillations: (a) general view; (b) schematic image; (1) plate; (2) corbel; (3) nut; (4) bracket; (5) coupling rod; (6) replacement beam; (7) pneumatic cylinder; (8) flange; (9) screw.",
        "texts": [
            " Original article submitted March 12, 2013; revision submitted October 17, 2013. 870 0021-8944/14/5505-0870 c\u00a9 2014 by Pleiades Publishing, Ltd. The experiments were performed in a T-313 supersonic wind tunnel based at the Khristianovich Institute of Theoretical and Applied Mechanics of the Siberian Branch of the Russian Academy of Sciences [5], which is a blowdown wind tunnel with the test section size of 0.6\u00d7 0.6 m. The model was tested on a dynamic setup with free rotational oscillations over the pitching angle with respect to the model-fitted coordinate axis OZ. Figure 1 shows the general view and the schematic image of the setup with the model, supporting devices, and actuation mechanism. A subsystem of oscillations with replacement beams is located at the end of a motionless sting. Together with the subsystem of oscillations, these beams define the offset (5.4 and 8.1 mm) of the axis of model oscillations with respect to the longitudinal axis. The subsystem of model fixation and release consists of a pneumatic cylinder attached at the tail part of the sting on a corbel",
            " The tests showed that the friction moment Mfr can be accurately described by the function Mfr = (\u2212Rf\u2217d/2) sgn (\u03b8\u0307), where R is the reaction force in the bearings, d is the bearing trunnion diameter, and \u03b8 is the pitching angle. As a result of calibration tests, the dry friction in the bearings could be reduced to f\u2217 = 0.001\u20130.002, which corresponds to the nominal data for the bearings used in this study. The pitching moment of the model in the wind tunnel was determined by processing the results of filming the model motion by a Phantom high-speed digital video camera through an IAB-451 shadowgraph. The reentry vehicle model was a capsule consisting of a spherical frontal shield and an inverse conical surface (see Fig. 1a). Two positions of the model rotation axis were considered: (1) xT = 88.8 mm and yT = \u22125.4 mm; (2) xT = 88.8 mm and yT = \u22128.1 mm. The coordinate xT was counted from the end section of the fuselage toward the model tip. The following geometric parameters of the model were used in calculating the aerodynamic characteristics of the model: reference area S = 0.0183 m2 and reference length l = 0.135 m. The moment of inertia of the model Iz with respect to the rotation axis was calculated by the \u201cSolid Works\u201d program"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002864_0954405416661003-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002864_0954405416661003-Figure8-1.png",
        "caption": "Figure 8. Manufactured two types of spiroid gear.",
        "texts": [],
        "surrounding_texts": [
            "Figures 9 and 10 illustrate the ease-off topographies of surface modifications and the results of TCA for case 1 and case 2, respectively. The fully conjugated surfacesPm 2 serve as the basis for calculating the desired ease offs. Figures 9(a-c1) and 10(a-c2) illustrate the synthetic ease-off topographies of the convex and concave tooth surfaces for two cases, respectively. It must be noted that ease-off topographies are conducted on the working surface due to the fillet surfaces generated by the generatrix of the addendum cylinder of virtual pinion hob that becomes impossible to implement surface modification. For the convex sides, the given bias angles h1 and h2 (shown in the Table 2) for the pre-designed paths of contact of two cases are not equivalent; the obtained ease offs of case 1 are smaller than that of case 2 at the corresponding grid points, as shown in Figures 9(a-c1) convex to 10(a-c2) convex, respectively. That means the modification amounts are influenced by the bias angle h of the desired contact path directly. However, regarding the concave surfaces, the bearing ratios of the two cases are designed to be different purposely, but the paths of contact are the same. Compared to case 1, the ease offs of case 2 are decreased, as shown in Figures 9(a-c1) concave and 10(a-c2) concave, respectively. In addition, if altering the assigned values of transmission error at control point (A\u2013D), the ease off along the at CORNELL UNIV on September 12, 2016pib.sagepub.comDownloaded from contact path will also be changed. This conclusion may be demonstrated easily. Figures 9(b-c1) and 10(b-c2) show the bearing contact on both the tooth surfaces of gear, respectively. The parabolic curve of the actual transmission error is provided by the actual rotation angles f1 and f2 of meshing gear pair. Three periods of cycle of meshing are considered to comparably analysis the conditions of meshing for the two examples, as shown in Figures 9(c-c1) and 10(c-c2), respectively. With respect to the TCA of the convex tooth surfaces, the results show that the magnitudes of transmission error of a cycle are 5 and 6.1 arc seconds, respectively. The contact ratio determined by the bias angle h of case 1 is about 1.5, compared to the 1.4 of case 2, which demonstrates the bearing capacity and the stability of gear 1 is higher than that of gear 2. About the TCA of the concave surfaces, the actual transmission errors Du2 are shown in Figures 7(c-c1) concave and 8(c-c2) concave, respectively. Comparing with the given magnitudes of transmission error, their magnitudes of a cycle of meshing are 6.1 arc seconds. Corresponding to the two gear drives, the contact ratio of gear drive 1 is 1.68, and the other is about 1.6, respectively. Using the results discussed above, the contact ratio of gear drive is not equivalent when the concave side or the convex side is chosen as the driving surface. Obviously, the contact ratio of the former is higher than the latter. This advantage makes the concave surface of gear as driving side of the helicon gearing in general. The TCA of computer simulation illustrates the bearing contact on both the gear tooth surfaces, which means point contact is achieved by the modification of the gear tooth surfaces."
        ]
    },
    {
        "image_filename": "designv11_64_0000395_978-94-007-4132-4_8-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000395_978-94-007-4132-4_8-Figure2-1.png",
        "caption": "Fig. 2. Centric angle-slider with conchoidal coupler curves (a) and a centric inverted slidercrank (b) which approximately replaces (a)",
        "texts": [
            " There is an additional list of 26 Hoecken models, partly belonging to the Reuleaux collection and partly distributed among chairs at the polytechnics in Karlsruhe, Danzig and Munich. In the following three papers of Hoecken are presented and discussed to give an idea of his mathematical skills, knowledge of kinematics and sense for practical applications. Conchoidal straight-line mechanism From the kinematic point of view the conchoidal straight-line mechanism is a fourlink centric angle-slider mechanism (Fig. 2a). The coupler link as the straight line gg constantly slides through the fixed sliding point O, whereas the movable sliding point A is led along the fixed straight line through Q having the distance a from O. The sliding direction of point A is perpendicular to the symmetry axis passing through O and Q, the latter being also the horizontal axis x of a rectangular coordinate system. Every point P on the coupler straight line gg follows a conchoidal curve. In case P is located on the opposite side of OA, the conchoidal curve contains a closed crunode. If this crunode is partly similar to a circle, it may be approximately replaced by a circle arc with centre point M on the x-axis and radius r (Fig. 2b). Thus a centric inverted slider-crank mechanism is found as a substitute mechanism, where the slider in A can be dropped, because its path partly consists of an approximately straight line, perpendicular to the x-axis [3]. Hoecken equates the polar equation of the crunode with that of a circle around M and takes into account that all coupler curves of the inverted slider-crank are symmetric related to the x-axis. He determines the dimensions of this substitute mechanism which approximately leads the coupler point A along a straight line through six points in finite neighbouring order with a distance of h/5 between two points each",
            " Moreover, Hoecken built two compact models in order to show the usefulness of his ideas and the validity of his calculations (Fig. 3). One should keep in mind that the mathematical aids at the time of Hoecken were limited and mainly consisted of tables of logarithmic and trigonometric functions. Slide rules were also in use, but not very efficient and/or precise. Solving Eqn. 1 for example, i.e. 0)cos(2)(cos)2()(cos2 222223 =\u2212\u22c5+\u22c5\u2212++\u2212\u22c5 aalreaelel \u03d5\u03d5\u03d5 (1) to calculate e, l, r, where three angles \u03c6 are given above or below the x-axis (cf. Fig. 2b), requires multiple interpolation and iteration procedures to gain results with four digits after the decimal point. The author used the commercial program \u201cMathcad 11\u201d and followed a sequence of equations published in [4]. Thus he could show that Hoecken\u00b4 calculations were correct (Fig. 4). Six-link dwell mechanism With dwell mechanisms having one d.o.f. the driven link temporarily comes to a standstill, while the driving link moves continuously. The standstill or dwell position is characterized by zero values of the velocity and acceleration of the driven link, related to the fixed link"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003109_iciea.2016.7603844-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003109_iciea.2016.7603844-Figure1-1.png",
        "caption": "Fig. 1. Prototype of the FSCW IPMSM",
        "texts": [
            " The harmonics of magnetic flux density in air-gap in noload and in load are analyzed in the paper. The harmonics contain slot harmonics, spatial harmonics and temporal harmonics of MMF. The influence on rotor magnet loss with different air-gap sizes is investigated by FEM, results show that widening the air-gap size can reduce the magnet loss. Both circumferential and radial PM segmentation are validated to reduce the magnet loss. II. ANALYSIS OF HARMONICS The prototype of an interior permanent magnet synchronous motor (IPMSM) with FSCW is shown as Fig.1. The main characteristics of the motor are contained in Table I . TABLE I. Motor main parameters PARAMETER Number of stator slots Number of rotor poles Rated Speed Peak Speed Rated Voltage Peak Power Peak Current Magnet residual induction(Br) at 25 Temperature coefficient of Br 24 16 1250rpm 5500rpm 245V 20KW 260A 1.26T -0.13%/ 1616978-1-4673-8644-9/16/$31.00 c\u00a92016 IEEE Fig. 2 shows the average eddy current loss in a magnet with different output power when the speed is 1250 rpm. When the output power is 15kw, the magnet loss is about 77"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000725_nme.4707-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000725_nme.4707-Figure1-1.png",
        "caption": "Figure 1. Geometric representation: (a) principal space and the two groups of base vectors and (b) geometrical relation between the normalized elastic strain deviator c, tensor d and the normalized stress deviator s, tensor t in the \u00a0 plane.",
        "texts": [
            " a1 C 2a2 a3/ (6) Then i ;x, y form another set of mutually orthogonal unit base tensors. This set of bases was used by Laine et al. [9] to formulate elastic constitutive equations. Using (6) and (5), one obtains the relations x D s cos \u2122 t sin \u2122; y D s sin \u2122C t cos \u2122 (7) Equations (5)\u2013(7) show that the transformation between the three sets of bases i ; s; t and i ;x;y as well as the eigenvalue bases ai .i D 1; 2; 3/ follows the transformation rule of vectors. They can be geometrically illustrated in the three-dimensional principal space, as depicted in Figure 1. As shown in Figure 1, i ;x;y constitute a global fixed coordinate system, and i ; s; t can be rotated from i ;x;y about the hydrostatic pressure axis by the Lode angle. Therefore, i ; s; t constitute a local cylindrical coordinate system, and the Lode angle plays the role of the polar angle. According to the geometrical interpretation, any arbitrary second-order tensor-valued function of the stress can be regarded as a vector in the principal space and the relations between its components associated with any two different sets of bases are established by the transformation rule of vectors",
            " Therefore, the base tensors c and d can be represented as the vectors in the Copyright \u00a9 2014 John Wiley & Sons, Ltd. Int. J. Numer. Meth. Engng 2014; 99:654\u2013681 DOI: 10.1002/nme principal space defined by the eigenvalue bases of the stress. With the geometrical interpretation of the Lode angle, the relationship between the two sets of base tensors can then be obtained by using the transformation rule of vectors \u00b9E\u00a9e\u00ba D \u0152R \u00b9E\u00a2\u00ba (14a) where [R] is the matrix describing counterclockwise rotation of an angle of \u00ae \u2122 about axis i , as shown in Figure 1b \u0152R D 2 4 cos.\u2122 \u00ae/ sin.\u2122 \u00ae/ 0 sin.\u2122 \u00ae/ cos.\u2122 \u00ae/ 0 0 0 1 3 5 (14b) It is noted that [R] is an orthogonal transformation matrix with its inverse equal to its transpose. The constitutive equations are formulated within the framework of small deformation, rateindependent elastoplasticity. The material behavior is assumed to be isotropic with a set of scalar internal variables characterizing the history of the deformation. For small strains, the total strain tensor is additively decomposed into its elastic and plastic parts \u00a9 D \u00a9e C \u00a9p (15) For isothermal deformation, standard thermodynamic arguments lead to the following constitutive relation \u00a2 D @W @\u00a9e (16) for the stored strain energy function W in terms of the elastic strain \u00a9e and a set of scalar internal variables, symbolically denoted by \u0178 D \u0178\u2019"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001186_icumt.2014.7002105-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001186_icumt.2014.7002105-Figure3-1.png",
        "caption": "Fig. 3. Track and bearing angles.",
        "texts": [
            " L2 and \u03bb2 are the longitude and latitude coordinates of the target waypoint. By using the two equations above the great earth circle distance between the two combined waypoints is calculated as shown below. . dist12 = tan\u22121 ( |P1 \u00d7 P2| P1 \u00b7 P2 ) (3) The bearing angle between the two pairs of coordinates is calculated by using the formula below. \u03b7P1P2 = P1 \u00d7 P2 |P1 \u00d7 P2| \u03b7P2P1 = P2 \u00d7 P1 |P2 \u00d7 P1| \u03c812 (4) = tan\u22121 ( \u2212\u03b7P1P2 P1\u03b7P1P2 \u2212 P1\u03b7P1P2 ) (5) The great circle distance and course angle of the flight route are shown in Fig. 3. The second common function used for the lateral navigation algorithm is the position calculation function. This function finds the longitude and latitude position coordinate values of the target waypoint from a given position vector, great circle distance and course angle values. For calculating the target waypoints longitude and latitude coordinates, primarily tangent unit vector in the course angle direction to great earth circle route is found. L1 is the latitude coordinate value of the first waypoint, \u03bb1 is the longitude coordinate value of the first waypoint",
            " L0 = tan\u22121 ( P01(3), \u221a 1\u2212 (P01(3)) 2 ) (11) \u03bb0 = tan\u22121 (P01(2), P01(1)) (12) If a parallel deviation from the flight route is required, the pitch and bank angle correction commands are calculated for fitting the desired flight path. Moreover, when the flight route switching waypoints are reached, these correction angles are calculated again in for introducing the next flight leg. [8] In order to calculate the lateral guidance command, firstly the projection waypoint of the current position of the air vehicle on the desired route is found. KTCP distance shown in Fig. 3, is the distance between the projection waypoint of the current position of the air vehicle and the point that the air vehicle will pass on the route. The position coordinate values of the point that the air vehicle will pass on the flight leg is found by the formula 7. The course angle between the current position vector of the air vehicle and the point RP is found by the formula 4. This course angle is called the desired track angle. ATD is the along track distance, the distance between the vertical projection waypoint of the current position of the air vehicle and the waypoint to be reached"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002031_0142331216631190-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002031_0142331216631190-Figure13-1.png",
        "caption": "Figure 13. Three-dimensional geometry of missile and target in beam-riding guidance law.",
        "texts": [
            " Performance of dither signals in different scenarios As mentioned in the previous subsection, the triangular signal has more suitable linear behaviour. Also, it causes a smooth shape of the angle of attack and sideslip, as shown in Figure 10. In this subsection, applying the beam-riding guidance law at MCMASTER UNIV LIBRARY on February 27, 2016tim.sagepub.comDownloaded from for two scenarios, the performance of different types of dithers is verified through simulations. The guidance law states that a beam (i.e. radar or laser) is continuously pointed at a target while the missile flies along it (Zarchan, 2001). Figure 13 shows the three-dimensional beam-riding geometry of the missile and target. The missile tries to decrease its distance from the centre of beam. The distance vector according to Euler angles u and c can be expressed as ~d u,c\u00f0 \u00de= RM uT uM\u00f0 \u00de RM cT cM\u00f0 \u00de cos uT\u00bd \u00f024\u00de Figure 14 displays a beam-riding guidance and control block diagram for missile and target engagement. at MCMASTER UNIV LIBRARY on February 27, 2016tim.sagepub.comDownloaded from Assuming Ad = 0:8 and fd = 55Hz for the two scenarios presented in Table 3, Figures 15 and 16 illustrate the distance of the missile from centre of beam along the flight for the two scenarios"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001451_ijrapidm.2014.066010-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001451_ijrapidm.2014.066010-Figure4-1.png",
        "caption": "Figure 4 Thermal distribution at the end of the exposure (see online version for colours)",
        "texts": [
            "15 mm) of the geometry chosen for this study is presented in Figure 3(b). On the model top surface, eight points are selected in such a way that these points are exactly in between two hatch lines where there is overlapping of the laser spot as depicted in Figure 3(a). Four simulation trials have been carried out by varying process parameters at four levels as listed in Table 2 and the resulting temperature values on the designated positions are recorded. The final temperature distribution on the top surface of the layer at the end of the exposure is shown in Figure 4. The maximum temperature is observed in the middle of the layer and the temperature falls gradually towards the edges of the section due to cooling from the sides. The maximum temperature on the layer reaches to 2500 K due to high energy density. This is mainly attributed to the high laser power and slow scanning speed (Figure 5). The stepping effect found on these temperature curves are the result of the second overlying exposures from the adjacent scans. Figure 6 presents the transient temperature response for different energy density values"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000174_978-0-387-92897-5_375-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000174_978-0-387-92897-5_375-Figure10-1.png",
        "caption": "Fig. 10 A typical averaged vertical strain time history",
        "texts": [
            " From these three expressions, a general bearing film stiffness law is found that applies for all restrictor-controlled recessed pads. l \u00bc 3prAe h 1 pr q : dq dpr h i \u00f01\u00de Stiffness depends on the restrictor flow relationships as stated in Figs. 2\u20134. Thus, If pr q : dq dpr < 1 the bearing is stable and stiffness is positive. If pr q : dq dpr \u00bc 1 the bearing has infinite static stiffness. If pr q : dq dpr > 1 the bearing is statically unstable since stiffness is negative. Flow characteristics of the four main types of control are shown in Fig. 10. Capillary restrictors give the lowest stiffness but tolerate small temperature variations better than other types of control because their resistance ratio is not a function of viscosity. Orifice restrictors give slightly higher stiffness values than capillary restrictors and constant flow supply even higher, but both are affected by changes in viscosity. A pressure-sensing valve tuned to give a flow-pressure slope of +1 gives infinite static stiffness. In practice, it is usual to accept slightly lower stiffness than infinite to ensure stability and avoid the possibility of limit-cycle oscillations",
            " It is generally accepted that no detectable yield stress as measured by the MRV is required for safe operation down to the \u201cW\u201d grade limiting temperature. In the MRV test configured for ASTM D4684 (ASTM D4684 2008), cooling is applied to a set of cells with the oil pre-charged in a slow-cooling profile designed to maximize any gelation tendencies in the fluid. Once the test temperature is reached, each oil sample is checked first for the presence of yield stress by placing progressively increasing 10 g weights on the string, until rotation is initiated (see instrument schematic, Fig. 10). ube Gallery b Before Pump Time P re ss ur e ) Flow-limited pumpability failure as seen in the oil pan (a) and AIR-BINDING PUMPABILITY CONDITION Critical Flow Path: Oil Surface to Inlet Screen Gallery ba Before Pump Time P re ss ur e Rheological Measurement Methods and Equipment, Fig. 9 (a, b) Air-binding pumpability failure as seen in the oil pan (a) and as detected by oil pressure in the oil circuit (b) R Shear stress at the rotor surface for this instrument is: tr \u00bc Tr 2pR2 r h 109 \u00bc 3:5M \u00f010\u00de where tr = Shear stress at the rotor surface, Pa Tr = Applied torque, N-m Rr = Rotor radius, mm h = Rotor height, mm M = Applied mass, g Thus, yield stress is tested in 35 Pa increments",
            " The P mP a Rigid Punch Problem with a Crack, Fig. 6 Circular-ended inclined punch y x P mP E a\u2032 Rigid Punch Problem with a Crack, Fig. 7 Flat-ended inclined punch with incomplete contact restraining condition for the smooth contact at the end E (s \u00bc b) is i\u00f01 im\u00de\u00f01 a\u00de 2pw\u00f01\u00de\u00f0b a\u00de P \u00fe \u00f01 m\u00def \u00f0b\u00de \u00bc 0 \u00f015\u00de Rigid Punch Problem with a Crack, Fig. 8 Wedge-ended punch with incomplete contact P mP a\u2032 Rigid Punch Problem with a Crack, Fig. 9 Circular-ended inclined punch with incomplete contact P mP Rigid Punch Problem with a Crack, Fig. 10 Circular-ended vertical punch with incomplete contact R The function f \u00f0s\u00de in (15) is given by f \u00f0s\u00de \u00bc G R \u00f01 im\u00de k\u00fe 1 E2 0 m 1 a \u00fe 1 m 1 b \u00fe 2E0Ec 1 w\u00f01\u00de\u00f01 s\u00de \u00fe E2 0 w\u00f01\u00de\u00f01 s\u00de2 \u00fe XN k\u00bc1 2E0Ek zk 1 1 w\u00f01\u00de\u00f01 s\u00de \u00fe 1 w\u00f0zk\u00de\u00f0zk s\u00de \u00fe XN k 6\u00bc1 EkE1 zk z1 1 w\u00f0z1\u00de\u00f0z1 s\u00de 1 w\u00f0zk\u00de\u00f0zk s\u00de \u00fe XN k\u00bcl E2 k w\u00f0zk\u00de\u00f0zk s\u00de2 \u00fe XN k\u00bcl E2 k m zk a \u00fe 1 m zk b \u00fe 2EcEk 1 w\u00f0zk\u00de\u00f0zk s\u00de 2 6666666666666664 3 7777777777777775 \u00f016\u00de The contact length a0 is obtained, a0 \u00bc o\u00f0a\u00de o\u00f0b\u00de \u00f017\u00de 1. Flat-ended inclined punch with incomplete contact (Fig",
            " The vertical force is applied on the y axis (e = 0). 3. Circular-ended inclined punch with incomplete con- tact (Fig. 9) P mP The stress function is presented by (2) \u00f0R 6\u00bc 1; e 6\u00bc 0\u00de. The vertical force is applied on the y axis (e = 0). The inclined angle e is determined by Rm \u00bc 0 in (14). Also the coordinate b of the smooth end E must be determined, satisfying Eq. (15) by the iterative calculation. The contact length is calculated by (17) (Hasebe and Qian 1997, 1998). E D a\u2032 4. Circular-ended punch not inclined with incomplete contact (Fig. 10) Rigid Punch Problem with a Crack, Fig. 11 Circular-ended vertical punch with incomplete contact at both ends The stress function is presented by (2) with H2\u00f0z\u00de \u00bc 0\u00f0R 6\u00bc 1; e \u00bc 0\u00de. The position of the punch e is given by (13) and (14). Also the coordinate b of the smooth end E must be determined, satisfying (15) by the iterative calculation. The contact length is calculated by (17) (Hasebe and Qian 1997, 1998). Shown in Fig. 11, both ends of the punch contact incompletely to the matrix. At these punch ends D and E, the stresses sr \u00bc try \u00bc 0, that is, the ends are in smooth contact and are not singular points",
            " 1982) Center shaft Specimen Specimen Load Spherical bearing Rolling Bearing Test, Fig. 8 Schematic construction of the essential part of the rig (Tokuda et al. 1982) LOAD DUE TO SPRING LOADING ASSEMBLY FLUID DRIP OR FLOW FEED 4 - THERMO COUPLE HOLES DRIVEN END BRG.1 BRG.2 OUTLET BRG.3 BRG.4 Rolling Bearing Test, Fig. 9 A section of the four-bearing fatigue life test rig (Hobbs 1982) SPLASH GUARD TEST LUBRICANT BALL RACE THRUST BEARING LOAD SPACER CHUCK (FITS INTO ROTATING SPINDLE) Rolling Bearing Test, Fig. 10 Rolling four-ball test layout (Eastaugh 1982) R A test rig contains four size 6,208 bearings. This rig was developed to provide a test method for lubricating fluids and their effect on fatigue life. Normally, 32 bearings are used for the life test to minimize the effect of scatter inherent in fatigue life testing. The normal rotational test speed is 3,000 RPM. The maximum Hertzian contact stress is 2.9 GPa. The failure is detected by accelerometers attached on the outer houses. Figure 9 shows the section of the four-bearing fatigue life test rig (Hobbs 1982). This test rig was developed in the early 1950s and has been widely used. It consists of four 12.7 mm diameter balls in a pyramidal arrangement to simulate an angular contact ball bearing operation, shown in Fig. 10. The upper ball is the test element with the rotational speed of 1,500 RPM. The remaining three balls are driven by the upper ball. The test rig was used to show that different lubricant-material combinations generated wide variations in the bearing performance (Eastaugh 1982). Normally, 600 kg thrust load is applied to the assembly. Most of the element RCF testers use the rolling motion to evaluate the stresses on each contact. However, the most important stresses influencing bearing life are the shear stresses occurring at EHL contacts",
            " 9 False flange development on a wheel by wear Head checks are surface defects and form like small hairline cracks on the railhead, as seen in Fig. 4. They are usually found at the wheel/rail contact locations. On straight track and in curves with gentle curvature, HC occurs at the top of the railhead, and for sharper curves, HC develops at the gauge corner. In gauge corner cracking, cracks develop at the corner of the rail. They are open to the surface. Cracks transverse to the rail direction can potentially cause a rail break. The appearance of GCC on the rail is shown in Figs. 5 and 6. Rolling Contact Fatigue (RCF), Fig. 10 False flange damage on rail Rolling Contact Fatigue (RCF), Fig. 11 Typical wheelburn R Rail squats are cracks that grow horizontally below the running surface and are thought to be caused by heavy wheel load over a cracked rail. This kind of RCF defect looks like a depression mark on the railhead (Fig. 7). The stresses due to trains passing over a rail deform the material and extrude it into a tongue, as shown in Fig. 8. During this process cracks initiate and develop at the interface between extruded material and non-extruded material. A new wheel has a conical tread and a flange on the inner side, as shown in Fig. 9a. After some period of operation, the wheel wears and a hollow may develop on the tread. Toomuch hollowing (because of excessive wear) can cause a false flange to develop on the outer side of the wheel (Fig. 9b). This damages the railhead, as seen in Fig. 10. Wheelburns result from frictional heating produced by a wheel spinning at the same spot on the rail head. Rapid heating followed by subsequent rapid cooling can cause damage at this spot and cracks may develop in the damaged material. The damaged material is highly brittle and can be spalled easily. A typical wheelburn is shown in Fig. 11. Tache ovale is a subsurface defect inside the railhead and has a kidney shape, as shown in Fig. 12. This kind of RCF is caused by manufacturing defects, such as hydrogen shatter cracks (small cracks), inclusions (undesirable elements or impurities), and voids (small holes)",
            " As the project was focused on the response of IRJ to heavy axle loaded traffic, strategies were required to sort out the data corresponding to the loaded trains. Several strategies were formulated based on the information on train composition, traffic data, and the strain signatures. First as the wheels do not run symmetric to the vertical axis of the rail, the strain components measured from the opposite faces of the rail web have been averaged. A typical vertical strain signature is shown in Fig. 10. Each spike corresponds to the averaged vertical strain at one of the rosette locations for 10 s of train travel (200,000 data points). The average strains were on the order of 500 microstrain. Each spike in the figure corresponds to one wheel passage across the joint (also the gauge location). Themagnitude of the strain values provide an indication if the wagon was loaded or empty (or partly filled). Furthermore, using the details of the composition of the trains and the design of bogies (e.g., axle spacing), it was possible to assess the speed of travel of the wheels",
            " The effects of the asperity contact have been accounted for using two approaches. In the first approach, the elasticity equations are numerically solved for the contact pressure and the shear deformation, in regions where contact occurs (Shi and Salant 2000). In the second approach, the contacts are treated as Hertzian contacts. In both approaches, the shear stress in contacting regions is 0.1 y* h 0 * 1.00.90.80.70.60.50.40.30.20.0 0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 V* = .0020 V* = .0025 V* = .0030 V* = .0035 V* = .0040 V* = .0045 Rotary Lip Seal Analysis, Fig. 10 Dimensionless film thickness distribution (Salant and Flaherty 1995) 0.0 0.1 0.2 5 6 4 2 3 0.3 0.4 0.5 \u03b4 * y* 0.6 0.7 0.8 0.9 1.0 5 y -oriented asperities 5 circular asperities 5 x -oriented asperities Rotary Lip Seal Analysis, Fig. 11 Dimensionless circumferential displacement distribution (Salant and Flaherty 1995) R determined using an empirical dry friction coefficient (Shen and Salant 2006). Due to normal engineering tolerances, the lips seal is never perfectly concentric with the shaft, nor are the lip and shaft surfaces perfectly axisymmetric",
            "025 or so for this set of analyzed cases. This is due to vanishing hydrodynamics at tiny film thickness ratios/average gaps. The mixed lubrication, where the contact load ratio is roughly in a range of 0 < Wc < 0.9, corresponds to the l ratio approximately between 0.025 and 1.5 for the analyzed line contact cases. In order to better compare the deterministic simulation results with those from well-known stochastic model by Patir and Cheng, line contact deterministic analysis results are re-plotted in Fig. 10. Although the same general trend is found in both the deterministic and stochastic analyses for the same parameter range of L = hcs/s > 0.5, the roughness orientation effect predicted by the deterministic model appears to be quantitatively less significant than that by the stochastic model. This is likely because the average flow factors used in the stochastic analyses did not consider roughness height reduction due to asperity deformation, so that the roughness effect on lubrication was overestimated. Also, the stochastic results were limited to relatively thick film cases, L > 0.5. When the speed approaches zero, L = hcs/s approaches zero in the meantime, but the decrease of l = ha /s is much slower due to the existence of rough asperities that are difficult to be completely flattened. That is why in Fig. 10 the ratio of ha/hcs may increase as the L ratio goes below 0.5. The situation for point contact cases may become more complicated. When the contact ellipticity, k = b/a, is large, the situation is close to that of line contact due to relatively insignificant lateral flows in the EHL conjunction. The basic trend for the surface roughness orientation effect may be similar to those of line contact, as shown in Fig. 8 for the cases of k = 2.0. If the contact ellipticity is small, e.g., k = 0.5, the lateral flows become significant and the entraining action is much weakened",
            " The roughness orientation effect on film thickness, therefore, appears to be less significant under the same operating conditions, but the longitudinal roughness is still relatively more favorable for lubrication formation in the analyzed cases. It is of extreme importance inmachine element design and production how to select, optimize, and specify surface finish. Today, requirements for continuously improving Line Contact Isotropic Transverse Longitudinal 10.1 h a /h s 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 \u039b= hcs/\u03c3 Patir & Cheng Roughness Effect on Elastohydrodynamic Lubrication, Fig. 10 Effect of roughness orientation on line contact EHL film thickness performance, efficiency, and durability, as well as reducing production costs, constantly impose great challenges to engineers and researchers. Since the effect of surface roughness and topography on the tribological performance and life is a complicated issue in practice, one or two statistic parameters conventionally used, such as Ra and Rq, are often found to be insufficient to satisfactorily specify the optimized surface finish"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003263_iros.2016.7759567-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003263_iros.2016.7759567-Figure1-1.png",
        "caption": "Fig. 1. Characterization of a snake robot into a single articulated-body. The handle is the link exerting the majority of contact wrench onto the object.",
        "texts": [
            " Therefore, it seems more natural that they grasp an object using their own body, in other words, an enveloping grasp [23] [24]. However, we propose that one of the links can be used to exert majority of contact wrench, and other links are used to keep the object grasped w.r.t. to the snake robot. This particular link can be thought of as the palm of a robotic hand. In this paper, we propose to simplify a snake robot as an articulated-body (AB) [25] and treat this particular link as a handle and project the equations of motion (n-dimensions) into this link\u2019s subspace (c.f. Fig. 1). The acceleration of link ai, considering the eqns. of motion (6), can be written as ai = \u03a6A i f \u2217 i + b A i , (8) where the terms \u03a6A i = J iM \u22121 s J i T bAi = J iM \u22121 s (B\u03c4 \u2217act \u2212 h\u2217s) + J\u0307 iq\u0307s denote the AB inverse inertia, and AB bias terms, respectively. The AB inverse inertia \u03a6A i is a mapping (positive semidefinite) from an external force applied to the link f\u2217i to its acceleration ai taking into account the configuration of the system. Assuming J i is full-rank, \u03a6A i is positive definite and therefore, invertible",
            " In other words, the correct expression is f\u2217c T \u03a6A i f \u2217 a = 0 (35) In this paper, we propose a simple parameterization to analyze the effect of the snake robot configuration on f\u2217c and f\u2217a , in order to deal with snake robots regardless of the number of joints. First, we propose to measure the distance from the COM of the whole snake robot COM s to the contact point c ds = ||COM s \u2212 c||. (36) Secondly, we use the average angle of the links of the snake robot ave\u03b8. Both quantities expressed w.r.t. the frame attached to the handle (c.f. Fig. 1). Four metrics based on the previous geometric analysis, are chosen to measure the performance of the snake robot. We concentrate on the maximum wrench that the snake robot can exert onto the object, so we ignore for now the velocity products of both systems and consider only the input joint torques: 1) The square of the magnitude of the contact wrench as a function of the input torques: \u03c81 := ||f\u2217c ||2 = f\u2217c \u25e6 f\u2217c = f\u2217cT\u03a6A i f \u2217 c = \u03c4 \u2217act TBTM\u22121JTi T\u0302 J iM \u22121B\u03c4 \u2217act, (37) where T\u0302 = T (T T\u03a6A i T ) \u22121T T ",
            " However, they are a little cumbersome when trying to compare configurations. To compare the performance of the snake robot in different configurations we use the maximum value of the metrics (e.g., max(\u03c81) for a given configuration). In order to apply the previous analysis, we study a snake robot with the parameters described in Table I. We assume the snake robot is contacting an object (fixed to the environment) with its tail (first link), and the contact occurs at the middle of the link (c.f. Fig. 1). This link is treated as the handle of the articulated-body. Finally, we assume the there is no friction at the contact. We change the angles of the joints in the range [\u221290\u25e6, 90\u25e6] every 5\u25e6 (resulting in 1369 configurations) and calculate metrics (37)-(40). Some configurations chosen in a fixed- interval can be seen in Fig. 2. The duality between f\u2217c and f\u2217a can be clearly seen by comparing the ellipsoid corresponding to \u03c81 and \u03c82. Fig. 3(a) and 3(b) show the value of the metrics as a function of the distance ds and the angle of the snake robot ave\u03b8, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.3-1.png",
        "caption": "Fig. 3.3 Partially closed kinematic chains",
        "texts": [
            "1007/978-3-540-36045-2_3, Springer-Verlag Berlin Heidelberg 2014 43 Closed kinematic chains\u2014kinematic loops Assuming an open kinematic chain with tree structure, a single independent kinematic loop is obtained by introducing an additional joint in each case. Using Eq. (3.1), the number of independent loops nL in such a kinematic structure is given by (Fig. 3.2) nL \u00bc nG nB: \u00f03:2\u00de Furthermore, partially and completely closed kinematic chains can be differentiated. A system with kinematic loops forms a partially closed kinematic chain, when \u2022 single partial systems form open chains or \u2022 multiple closed partial systems are connected to each other in an open chain (Fig. 3.3). A chain can be considered completely closed when \u2022 each body is a part of a multibody loop and \u2022 each loop has at least one body that is connected to another loop. A mechanism, by definition, must be a partially or completely closed kinematic chain (for more details see Sect. 3.4.1) (Fig. 3.4). Kinematic chains can be grouped into three distinct motion categories: Planar kinematic chains In a planar kinematic chain all body points move inside or parallel to a reference motion plane (Fig. 3.5)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002824_chicc.2016.7554386-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002824_chicc.2016.7554386-Figure1-1.png",
        "caption": "Fig. 1: The torpedo-like AUV and the reference frames",
        "texts": [
            " M \u03bd\u0307 + C(\u03bd)\u03bd +D(\u03bd)\u03bd + g(\u03b7) = \u03c4 (1) where \u03bd and \u03b7 are velocity and position/attitude vectors, respectively; \u03c4 is the model control input; M , C(\u03bd), and D(\u03bd) are inertia, Coriolis and centripetal force, and damping matrices, respectively; g(\u03b7) is the force/torque vector produced by the gravity at the buoyant center. Formula (1) can be simply expressed as \u03bd\u0307 = fN(\u03bd,\u03b7) +M\u22121\u03c4 (2) where fN(\u03bd,\u03b7) is an abbreviation based on formula (1). The earth-fixed reference frame o-xyz, the body-fixed reference frame oB-xByBzB, and the torpedo-like AUV are illustrated in Fig. 1, which will be the research platforms in the following sections. The model control input \u03c4 in formula (1) is a resultant force/torque vector, which is an expression of effects brought by a group of actuators in general. For the torpedo-like AUV as shown in Fig. 1, \u03c4 is usually expressed in the form of formula (3) [14]. \u03c4 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 I 0 0 0 0 1 2 \u03c1Sv\u03032C\u03b4r y 0 0 0 0 1 2 \u03c1Sv\u03032C\u03b4e z 0 0 1 2 \u03c1SLv\u03032m\u03b4r x 0 1 2 \u03c1SLv\u03032m\u03b4d x 0 0 1 2 \u03c1SLv\u03032m\u03b4e y 0 0 1 2 \u03c1SLv\u03032m\u03b4r z 0 0 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u23a1 \u23a2\u23a3 T \u03b4r \u03b4e \u03b4d \u23a4 \u23a5\u23a6 (3) where \u03c1, v\u0303, S, L, and T are seawater density, AUV speed, maximum cross-sectional area of AUV, AUV length and propeller thrust, respectively; \u03b4r, \u03b4e and \u03b4d represent vertical, horizontal and differential rudder angles, respectively; C\u03b4r y and C\u03b4e z describe the position derivatives derived from the corresponding force factors by \u03b4r and \u03b4e separately; m\u03b4r x , m\u03b4d x , m\u03b4e y , and m\u03b4r z are position derivatives derived from the corresponding moment factors by the relevant rudder angles",
            " To tolerate the additional rolling torque, the other part of the differential rudder angle could be \u03b4\u2021d = \u2212 1 2\u03c1SLv\u0303 2m\u03b4r x (\u03b4r \u2212 \u03b4 r ) 1 2\u03c1SLv\u0303 2m\u03b4d x = 2m\u03b4r x f\u0302N \u03c1v\u03032SLm\u03b4r z m\u03b4d x (19) Thus, the designed differential rudder angle is \u03b4d = \u03b4\u2020d + \u03b4\u2021d = \u03b4 d \u2212 2f\u0302K \u03c1v\u03032SLm\u03b4d x + 2m\u03b4r x f\u0302N \u03c1v\u03032SLm\u03b4r z m\u03b4d x (20) Based on the above designing, the resultant torque of the rolling movement could be K = 1 2 \u03c1SLv\u03032m\u03b4r x \u03b4r + 1 2 \u03c1SLv\u03032m\u03b4d x \u03b4d + fK = 1 2 \u03c1SLv\u03032m\u03b4r x ( \u03b4 r \u2212 2f\u0302N \u03c1v\u03032SLm\u03b4r z ) + fK+ 1 2 \u03c1SLv\u03032m\u03b4d x ( \u03b4 d \u2212 2f\u0302K \u03c1v\u03032SLm\u03b4d x + 2m\u03b4r x f\u0302N \u03c1v\u03032SLm\u03b4r z m\u03b4d x ) \u22481 2 \u03c1SLv\u03032m\u03b4r x \u03b4 r + 1 2 \u03c1SLv\u03032m\u03b4d x \u03b4 d (21) which means the rolling torque approaches to the resultant torque produced by the desired vertical rudder angle and the desired differential rudder angle, and the fault effect generated by the horizontal rudder fault has been effectively tolerated. Through the above processes, the FTC is finally realized. The FTC algorithms for the other rudders are analogous to the aforementioned one, which are omitted here. Simulation results are presented in this section to illustrate the aforementioned methods. We take a torpedo-like AUV as shown in Fig. 1 to perform a depth control simulation. The AUV parameters can be got from the appendix of ref. [14]. The rudder deformation fault is designated at simulation time 77s on RR with the bent angle being 1.05rad, the \u2220A = 0.52rad, and the length of AB being 0.3m. Under the given deformation fault, the depth control result is displayed by the black dotted line as shown in Fig. 3, where the blue double dash line indicates the sailing result without any rudder fault. The vertical distance between the two lines manifests the influence on the depth control caused by the RR deformation fault"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002948_gt2016-57458-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002948_gt2016-57458-Figure4-1.png",
        "caption": "Figure 4 Problems related to the SLM processing of overhang sections",
        "texts": [
            " Main function of the DFP is to damp thermo-acoustic flame instabilities. Therefore, damper volumes are needed to reduce high frequency amplitudes of pulsations coming from the combustion chamber. These damper volumes were the first zones of high potential risk of overhangs for the SLM process. Overhangs in general are critical to the SLM process as the heat from the laser system is more and more transferred into the powder and not into solidified layers. This decreases the surface quality and accuracy and can cause failure for the manufacturing as illustrated in Figure 4. For a successful completion of the SLM process, it is necessary to avoid all the critical overhangs. Hence, the damper front panel was inclined with respect to the substrate plate to improve the SLM production. The chosen built-up orientation allowed minimizing the risk of overhanging walls. A second constrain, which had to be obeyed, was the maximum allowable built chamber size of the SLM machine. The final built-up orientation of the damper front panel allowed also positioning the part within the SLM machine chamber"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure10-1.png",
        "caption": "Fig 10 Dynamic eccentricity angle",
        "texts": [
            " Hence we should try to avoid large static eccentricity and make better efficiency and lower magnet loss. B. The influence of dynamic eccentricity on rotor loss In the case of dynamic eccentricity, the position of minimum air gap length rotates with the rotor position making the eddy current loss on magnet different with static eccentricity. The rotating axis is off-center from rotor axis. The effect of dynamic eccentricity varies with different eccentric angle. Eccentric angle is the included angle between rotating axis and rotor axis as Fig 10 show. So the effect of eccentric angle on magnet will be calculated on dynamic eccentricity. Since for most of the additional harmonic magnetic field, rotor slip is zero under dynamic eccentricity, so whether no-load or load will not cause large additional loss in the rotor. Healthy and faulty machines with different eccentric angle under different load conditions were simulated. The result is shown that dynamic eccentricity only increases a little rotor loss compared with healthy machine in Fig 11"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000627_dsn-w.2015.27-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000627_dsn-w.2015.27-Figure1-1.png",
        "caption": "Fig. 1: UAV Axes",
        "texts": [
            "27 118 1 4 73 8044 7 An airplane has six degrees of freedom (DoF), being three translation motion, i.e vertical, horizontal and transverse and 3 rotational motion, i.e, roll(\u03c6), pitch(\u03b8) and yaw(\u03c8) [11]. The traditional control surfaces on an aircraft are: \u2022 Ailerons: Controls the rolling. \u2022 Elevator: Controls the pitching. \u2022 Rudder: Controls the yawing. \u2022 Throttle: Controls the motor speed. The aircraft used in this work doesn\u2019t have all the traditional surfaces, instead it uses what is normally called elevons [12], these acts like ailerons and elevators at the same time, as shown in figure 1. The control of an aircraft is divided in lateral controller and longitudinal controller. 1) Lateral Controller: The lateral controller controls the roll angle and the heading of an aircraft by changing the angle of ailerons and rudder, or in this case the elevons. 2) Longitudinal Controller: The longitudinal controller controls the pitch angle, the airspeed and the altitude of an aircraft by changing the throttle and the elevator angle, again in this case the elevons. As noted above, both controller will change the angle of the elevons, so in a reduced control surfaces aircraft, the resulting controller will be a mix between the two controller",
            " The controller diagram is shown in figure 11, its inputs are the roll rate error \u03b5p, the yaw rate error \u03b5r, and their variations in time, assuming the roll rate p at a time t is represented as p(t), the error is given by equation 13 and the same is true for yaw rate r, equation 14. \u03b5p = p(t)\u2212 p(t\u2212 1) (13) \u03b5r = r(t)\u2212 r(t\u2212 1) (14) Roll Rate Error Rate and Yaw Rate Error Rate are the rate of changing in \u03b5p and \u03b5r. The aircraft used is a delta wing aircraft, commonly called combat wing, as seen in figure 1. This model is already implemented in the flight simulator XPlane 10, as seen in figure 12. B. XPlane 10 XPlane 10 is a flight simulator certified by FAA (Federal Aviation Administration). It was used to run the hardware in loop (HIL) simulations, a system where the autopilot hardware is connected to flight simulator in order to control the flight. Figure 13 shows the NavStik board, which is the smallest autopilot hardware available on the market, it features a 168MHz 32-bit ARM Cortex M4, capable of real-time processing, and 3 axis accelerometer, 3 axis gyroscope, 3 axis magnetometer, barometer, GPS and a Differential pressure sensor as navigation sensors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.12-1.png",
        "caption": "Fig. 4.12 Vector components in global and local systems of coordinates",
        "texts": [
            " However, it is worth noting that the absolute coordinates have the great merit to be quite straightforward, even for systems with high level of complexity. If a planar multibody system is made of nb rigid bodies such as the one illustrated in Fig. 4.11, then the number of absolute coordinates is n = 3 \u00d7 nb. Thus, the vector of generalized coordinates of this system can be written as q \u00bc qT1 qT2 . . . qTnb T \u00f04:22\u00de Let now consider a single body, denoted as body i, that is part of a multibody system, as Fig. 4.12 depicted. When absolute coordinates are used, the position and orientation of the body are defined by a set of translational and rotational coordinates. Thus, body i is uniquely located in the plane by specifying the global position, ri, of the body-fixed coordinate system origin, Oi, and the angle \u03d5i of rotation of this system of coordinates with respect to the x-axis of the global coordinate system. Hence, the vector of coordinates of the body i is denoted by qi \u00bc xi yi /if gT \u00f04:23\u00de Let ux and uy be absolute components of a vector u along the global x and y axes, respectively, and let u\u03bei and u\u03b7i be vectors along the body-fixed axes \u03bei and \u03b7i, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001972_ilt-03-2015-0034-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001972_ilt-03-2015-0034-Figure2-1.png",
        "caption": "Figure 2 Integration point within an element of a control volume",
        "texts": [
            " The equations governing the flow are presented in their averaged form of a Cartesian coordinate (x, y, z): Turbulent flow behavior in plain journal bearing Nadia Bendaoud, Mehala Kadda and Abdelkader Youcefi Industrial Lubrication and Tribology Volume 68 \u00b7 Number 1 \u00b7 2016 \u00b7 76\u201385 D ow nl oa de d by F lo ri da A tla nt ic U ni ve rs ity A t 2 2: 59 0 6 M ar ch 2 01 6 (P T ) Xj ( Uj) 0 (8) Xj ( UjUi) P Xi Xj Ui Xj Uj Xi Bx (9) Equations (8) and (9) can be integrated into a control volume, using the divergence theorem to convert Gaussian integrals into surface integrals volume as follows: S Uj dnj 0 (10) S UjUidnj S Pdnj S Ui Xj Uj Xi dnj V Suidv (11) The next step is to discretize the m known problem as differential operators of this equation. All these mathematical operations lead to getting on each volume control, a discretized equation linking variables from one cell to those of neighboring cells. All these discretized equations ultimately form a matrix system. Considering now an element of an isolated mesh such as that shown below in Figure 2. After discretization and rearrangement of equations (10) and (11), we will obtained the following forms: ip ( Uj nj)ip 0 (12) ip mip(Ui)ip ip (P nj)ip ip Ui Xj Uj Xi nj SuiV (13) Nj 1 i j 0 i j The interpolation method of the pressure in the pressure\u2013 velocity coupling is similar to that used by Rhie and Chow (1982). This method is among the best methods to save one memory space and computation time. If the pressure is known, the discretized equations are solved easily. The equation of mass conservation for a single dimension can be written as follows: U x i x3A 4m 4P x4 0 (14) Where: m Ui nj (15) The solution fields are stored in the mesh nodes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003424_ecce.2016.7854806-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003424_ecce.2016.7854806-Figure4-1.png",
        "caption": "Fig. 4: Comparison of the magnetic flux density between a healthy machine and a machine with 80% eccentricity.",
        "texts": [
            " 3b shows the average of the x and y component of the magnetic flux from all the poles under healthy and four severities of eccentricity fault. For healthy machine, the average x and y components equal to zero, but as the severity of eccentricity fault increases, the average of the x and y components will increase, this will bring up an increase in the total flux. It can also be noticed that the main increase was in the x axis because the shift of the rotor and the rotation axis was in the positive x direction Fig. 4 shows a comparison of the magnetic flux density between a healthy machine and a machine with 80% static eccentricity fault. The increase in the magnetic flux density implies an increase in the value of both \u03bbd and \u03bbq . Based on (4) and (5), increasing \u03bbd and \u03bbq , for the same operating load, will increase Vd but decrease Vq . So in the case of an eccentricity fault the value of the point (Vd, Vq) in the Vd-Vq plane will shift to the top left, and as the severity of eccentricity fault increases the point (Vd, Vq) will shift more"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000783_s13344-014-0065-9-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000783_s13344-014-0065-9-Figure1-1.png",
        "caption": "Fig. 1. Miniature autonomous underwater vehicle: (a) lateral view; (b) top view.",
        "texts": [
            " This vehicle has many advantages compared with regular AUVs because of its small size and low cost. Furthermore, the vehicle is designed to have three independent actuators in surge, heave, and yaw directions; consequently, it is more maneuverable than regular AUVs. This miniature AUV is about 40 cm in length and 10 cm in diameter. It has a cylindrical shape which is very typical for underwater vehicles. The head of the AUV is elliptical and the tail is conical. These elements are purposefully selected in order to reduce the drag forces (Fig. 1). As shown in Fig. 1, the propulsion is achieved by a central thruster. Also, two lateral waterproof motors with propellers are mounted just behind the centerline to provide yaw control and two motors are mounted vertically to control the depth of the vehicle. Roll and pitch are controlled by mounting the majority of the weight at the bottom of the AUV. Passively controlled, the vehicle can resist external disturbances. Moreover, because of the designed mechanisms for motions in heave and yaw directions, the vehicle does not need to have roll and pitch motions to go up and down and to turn left and right, so it always maintains in a horizontal posture"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001477_s00170-015-7015-4-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001477_s00170-015-7015-4-Figure12-1.png",
        "caption": "Fig. 12 Experimental coating workpiece Fig. 13 Coating bond strength test",
        "texts": [
            " The cutting residual compressive stress in the coating surface was larger than the tensile stress; therefore, the machined coating surface was under a compressive stress. In addition, as influenced by arc spraying, thermal residual stress, and cutting residual stress, the residual stress on the coating was tensile at depths below 0.2 mm. After the first cutting step (to 0.35-mm depth), the orthogonal test L9(3 4) was used to study the effect of cutting parameters on coating bond strength \u03c3 (see Fig. 12 and Table 3). In Table 3, \u03c3c=(\u03c3o\u2212\u03c3)=(50\u2212\u03c3), where \u03c3o is the original bond strength, and \u03c3c represents the effect of cutting parameters on the reduction in bond strength. To test the coating bond strength, a glue (type E-7) was chosen for these experiments: its bond strength can reach the 70 MPa after 3 h at 100 \u00b0C. Then, an MTS809 tensile testing machine was used (see Fig. 13). Based on the data in Table 3, \u03c3c was defined as: \u03c3c \u00bc Caxp f yvz; \u00f06\u00de which could also be expressed as: ln\u03c3c \u00bc lnC \u00fe xlnap \u00fe yln f \u00fe zlnv: \u00f07\u00de Then, the effect of cutting parameters on the bond strength is shown in Table 4 based on the analysis of variance (ANOVA), and the variables of Eqs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001742_s00542-015-2460-4-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001742_s00542-015-2460-4-Figure5-1.png",
        "caption": "Fig. 5 The application of the external shock using the large mass method",
        "texts": [
            " The fact that the vertical 1 3 force for these two cases was almost identical shows that the effect of the stiffness of the air bearing over a wide range was negligible as in a previous study (Zeng and Bogy 2002). The slider was connected to the disk using four linear springs, as shown in Fig. 4. These springs were located at the center of trailing edge (TE), the leading edge (LE), inner edge (IE) and outer edge (OE) of slider. When an external shock was applied, it was transmitted through four bolt points. The large mass method is one of methods to describe the external shock transmission, whereby a large mass was connected to the four bolt points as shown in Fig. 5. And, the external shock is applied to the large mass to investigate the shock response. With the contact model, it is necessary to establish which surfaces are contact and target surfaces. As shown in Fig. 6, both sides of the ramp were considered to be contact surfaces, and both sides of the disk were target surfaces. The augmented Lagrangian formulation includes controls to automatically reduce the penetration, and supports symmetric behavior. By exploiting this symmetric behavior, the contact and target surfaces were constrained from penetrating each other"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001786_vppc.2015.7352971-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001786_vppc.2015.7352971-Figure7-1.png",
        "caption": "Fig. 7. Models to confirm the effect of demagnetized pole arrangement",
        "texts": [
            " However, the distortion of input current waveform is not detected. Figure 6 shows the current spectrum density of each case. In Fig. 5, (b),(c) and (d) are different from normal condition (a). On the other hand, (e) and (f) is not different from (a). When odd numbers of poles are demagnetized, it can be detected by applying the proposed method. However, when even numbers of poles are demagnetized, the diagnosis result is not clear. Fig.6. Current spectrum density according to numbers of demagnetized poles Figure 7 shows the models to confirm the effect of demagnetized pole arrangement. In the Figure 7, (a) and (c) are asymmetrically demagnetized models. On the other hand, (b) and (d) are symmetrically demagnetized models. The axis of symmetry is indicated by using the dotted red line as shown in Fig. 7. In Fig. 8, (a), (b) and (c) are different from the normal condition, but (d) is not different. From this result, we found that this method cannot detect the demagnetization when the fault is periodically occured in each period of electric angle. III. CONCLUSION In this study, demagnetization diagnosis method based on the result of locked rotor test is proposed. The proposed method is capable of detecting the irreversible demagnetization without disassembly of motor. In order to verify the proposed demagnetization diagnosis method, each case is distinguished between the normal and fault model by using the FEA"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001553_2015-01-1567-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001553_2015-01-1567-Figure2-1.png",
        "caption": "Figure 2. 3-DOF Steering Model Schematic",
        "texts": [
            " One such example is shown in [9], where a controller was incorporated to illustrate how system gains influenced response, such as torque effort. The details of EPS were still limited, however. The aim of this paper is to develop a detailed steering model, with the goal of predicting on-center response (feel). A three DOF model will be presented and the implementation of a rack EPS will be included. The power steering boost curves (maps) will be shown and the model will be validated using test data. Details of how to tune the model will also be discussed. A schematic of the three DOF steering model used in this research is shown in Figure 2. It is coupled with a three DOF vehicle model and is similar to the one developed by [5] and is common in the literature. For brevity, we will focus on the steering model within. The steering model consists of two rotating masses (inertias): a steering wheel and shaft and an intermediate shaft; and one translating mass: the rack and effective masses of the pinion and motor. The upper shaft (steering wheel) and intermediate shaft are connected via a column (rag) coupler; and the pinion and intermediate shaft are connected through the torsion bar"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000318_tia.2010.2070784-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000318_tia.2010.2070784-Figure5-1.png",
        "caption": "Fig. 5. Twelve-slot ten-pole motor prototype.",
        "texts": [
            " 2) It is characterized by a high self-inductance, which is necessary to limit the short-circuit current [9]. 3) In addition, such a combination of the number of slots and poles yields a very low mutual coupling among the phases, avoiding mutual interaction of faulty phases and healthy phases [10]. Aside from the double-layer (DL) winding, the single-layer (SL) winding can be chosen so as to eliminate the physical contact between phases [2], [11]. A 12-slot 10-pole IPM motor prototype has been manufactured, as shown in Fig. 5. The nonoverlapped coils of the fractional-slot winding can be easily recognized. All the coils are separated, and the terminals of each coil are available (see Fig. 5), so that several winding connections can be tested, including both DL and SL. However, the SL winding is achieved by disconnecting every second coil. Therefore, the derived SL winding results to have a number of turns Nt that are half of the allowable rated number. The inverter that is available for the tests has six separate fullbridge converters. It is used for both three- and six-phase supplies, and its maximum phase current is equal to 6.2 A (peak). In this section, some considerations are carried out about the arrangement of the two sets of three-phase windings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003048_ebccsp.2016.7605267-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003048_ebccsp.2016.7605267-Figure2-1.png",
        "caption": "Fig. 2. Group of N -VTOL UAVs",
        "texts": [
            " the desired orientation, then the quaternion that represents the attitude error between the current orientation and the desired one is given by q\u0303 = q\u22121 d \u2297 q = (q\u03030 q\u0303 T v )T (5) where q\u22121 is the complementary rotation of the quaternion q which is given by q\u22121 = (q0 \u2212 qTv ) T and \u2297 denotes the quaternion multiplication [22]. In the case that the current quaternion and the desired one coincide, the quaternion error becomes q\u0303 = (\u00b11 0T )T . As it was mentioned before, the quaternion representation is redundant. As a consequence, the error mathematical model has two equilibrium and this fact must be considered in the stability analysis [23]. Now, consider a group of N -VTOL UAVs modeled as rigid bodies (see Fig. 2). Then according to the aforementioned and to [21], the six degrees of freedom model (position and attitude) of the system can be separated into translational and rotational motions, represented respectively by \u03a3Ti and \u03a3Ri in equation (6) and (7). \u03a3Ti : \u23a7\u23a8 \u23a9 p\u0307i = vi v\u0307i = g e f 3 \u2212 1 mhi RT i Ti e b 3i (6) \u03a3Ri : \u23a7\u23a8 \u23a9 q\u0307i = 1 2 \u039e(qi)\u03c9i Ji\u03c9\u0307i = \u2212[\u03c9\u00d7 i ]J\u03c9i + \u0393i (7) with i \u2208 N = {1, ..., N}. mhi denotes the mass of the ith VTOL-UAV and Ji its inertial matrix expressed in Eb i . g is the gravity acceleration and e b 3i = e f 3 = (0 0 1)T "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001380_sami.2015.7061901-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001380_sami.2015.7061901-Figure8-1.png",
        "caption": "Figure 8 Test plate \u201cTP01_SEC.01_P01\u201d in section with the upper hatching teeth detail",
        "texts": [
            " roughness), according to their mutual placement, with sufficient distance between the individual testing plates, or sectors, without a so-called \u201cinfluence zone\u201d. The arranging of the individual test plates into sectors labelled, i.e. TP01_SEC.01_P01 up through P05, TP01_SEC.02_..., TP01_SEC.03_..., TP01_SEC.04_..., TP01_SEC.05_..., TP01_SEC.06_..., TP01_SEC.07_..., TP01_SEC.08_..., TP01_SEC.09_..., is displayed in the following figure (Fig. 6, 7). E. Type of support material for a test plate Only one type of support material was used for all of the testing plates (i.e. from \u201cTP01_SEC.01_P01\u201d up through \u201cTP01_SEC.09_P05\u201d). Fig. 8 shows a sectioned view of a testing plate (specifically \u201cTP01_SEC.01_P01\u201d) together with a detailed view of the type of support material, or the characteristic shape of the so-called \u201cupper hatching teeth\u201d of the model/product. F. Manufactured test plates III. MEASURING OF SURFACE ROUGHNESS OF THE TEST PLATES IN THE INDIVIDUAL SECTORS This part of the work contains a description of the measurement of the surface roughness of the individual testing plates with respect to the sectors in which the testing plates were located"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure9-1.png",
        "caption": "Fig. 9. Force Diagrams for Brakes Using \u201cThree-to-one\u201d Sliding Shoes",
        "texts": [
            " F = (+-++1)0.75/2 = 0.75. -Reactions on carrier plate. - - - + Shoe reactions not acting directly on carrier plate. Drum reactions are not shown. total carrier plate reaction x k total shoe-tip load F = brake factor = movement of the shoe tips is restricted only by the friction of the pistons and rubber cups in the wheel cylinder. This frictional force and also the frictional forces at the sliding abutments are ignored in the present calculations. The force diagram for this type of brake is shown in Fig. 9a, from which it will be clear that the lining on the leading shoe can be expected to wear at least three times as quickly as that on the trailing shoe. This difference can be reduced by increasing the diameter of the piston which operates the trailing shoe, but this will reduce the brake factor because the leading shoe piston will have to be reduced in diameter in order to maintain the original total shoetip load. A brake of this type with equal shoe factors at p = 0.35, and therefore equal lining wear as long as this coefficient of friction is maintained, is shown in Fig. 9e and curves are shown in Fig, 10 for both straight and stepped cylinders, as well as for the other three types of brake. In the Girling brake an additional force is acting in the direction of the trailing-shoe tip, equal to the pull in the rod multiplied by the coefficient of friction between the expander housing and the carrier plate on which it is intended to float. With this type the contrast in wear between the leading and trailing shoe linings seems to be less marked than with the hydraulic brake, and this frictional force restraining the movement of the expander may be the reason",
            " These results are only true as an average during the life of the linings, the varying deflexions at different operating loads being likely to give instantaneous results which are quite unpredictable. It also usually happens that the cam imparts unequal movements, so that its direction of rotation relative to that of the drum is important, but this has been fully dealt with elsewhere (Dawtrey 1930) and the effect is omitted from the present calculations. Equal movements can also be obtained by intercoupling the shoes by a separate linkage and using a floating expander, and this method makes possible fixed cam characteristics with hydraulic operation (Fig. 9f). The diagram for the type is shown in Fig. 9b and the curve appears in Fig. 10. The low output and great stability compared with the identical brake with floating expander will be noted. With an undue rise in friction value it would presumably be possible for the leading shoe to become self-applying and to sprag by leaving the cam behind, unless the two shoes were positively interlinked, with a tension member to hold back the leading shoe, but in practice this seldom happens with a fixed cam brake. at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from (Reciprocal faAors of brakes with sliding shoes having parallel abut- ments.) I = 1.12 and K = 0.75. 1. Floating expander. 4. Duo-servo. 2. Fixed cam. 3. Two-leading shoe. 5. Stepped cylinder (3/1). Two-leading-shoe Brake. This brake has exactly the same characteristics as a single-leading shoe (Fig. 9c). The drum drag is doubled but twice as much aggregate shoe tip movement has to be imparted, so that the factor remains unchanged. When separate hydraulic cylinders are used (Fig. llu), the two shoes are fully compensated for both load and movement so that the a Single-acting hydraulic. BRAKES FOR ROAD VEHICLES 47 drum reactions are balanced and produce no off-set load on the wheel bearings, whilst the two linings will run at approximately the same temperature and wear equally, but with the mechanical type using bell crank levers and a strut (Fig",
            " A P P E N D I X I D E T E R A M I N A T I O N OF FORMULAE FOR SHOE FACTORS (as defined on pp. 44 and 45) ( 1 ) Factors for sliding shoes with parallel abutments. ON In Fig. 5, drum drag = R X -. r Shoe-tip load = L for leading or M for trailing shoe. R x O N o r R x O N Therefore F = - -L x r M x r - A B x O N o p A B x O N B N x r A N x r 2kl sine 2kl sin ek-1sinOor k + l s i n f l -- (2) Factors for brakes using sliding shoes with parallel abutments. The brake factor is the mean of the shoe factors. The brake layouts are shown in Fig. 9. (a) Floating expander :- 21 cosec 8 cosec2 e - 121k2\u2019 - (b) Fixed cam :- Let x = leading-shoe-tip load, and (2-x) = trailing-shoe-tip load; then (2-x) = leading-shoe-abutment load, and x = trailing-shoe-abutment load. In Fig. 9b, F = 2 { ( 2 - ~ ) - ~ ) k / 2 22 . F = cosec 0 - I/k 21 cosec 0-1/k\u2019 (d) Duo-servo :- Primary-shoe factor = Secondary-shoe-tip load = primary-shoe-abutment load k+lSinB k - I sin 8 --. 21 k+I sin0 heref fore F = t-(cosec e- l /k) ( l+- l 21 e - (cosec O--l/k)2\u2019 (e) Stepped cylinder, ratio s / l :- Let leading-shoe-tip load be x. Then trailing-shoe- tip load is ( s x x ) and X + ( S X X ) = 2. Therefore and 2 (S+l) 2s ( s x x ) = -- ( s+ l ) \u2019 2 21 2s x = - \u201c i = *{m) X(cosecB-Z/K)+(sx(cosec8+I/k) - -i (s + 1) 1",
            " One type of moulded lining which he had tried was very much more robust than most, and it would stand quite a lot of maltreatment before it began to crack. Mr. G. J. WARING (Leyland) agreed with the author that for practical purposes the minimum stopping distance of a vehicle was when the brakes locked all four wheels. A braking stop of such a description would automatically have the effect of eliminating brake squeal, fade, and temperature problems in the brake arrangement, but would transfer those common troubles to the tyres, which would then have a limited life. Theoretically the force diagrams in Fig. 9 were correct, but he wondered how brake drum distortion would affect those forces outlined in the diagram. Wheel bearing clearance might also have some effect on those problems. Investigation of squeal was often very dif\u20acidt owing to the unpredictability of its occurrence. Frequently a squealing brake when dismantled and then reassembled without modification would give no trouble, at least for a time, from squeal. He asked whether a road vehicle assembly that would squeal readily had been produced in the laboratory, and if that had been done, what the conditions of the lining and the drum were under squealing conditions",
            " WALLER wrote, in reply to questions on brake shoe layouts, that the reason for the diversity of types of brake used on road vehicles was that conflicting requirements had to be fulfilled and the compromise reached depended on the size and performance of the vehicle and whether or not power assistance was provided for the operating mechanism, while the interlinkage of the handbrake often resulted in the rear brakes being of a type different from the front. Table 2 showed representative figures comparing the suitabilities of the different types of brake in respect of three requirements, the conflicting nature of which would be seen :- (1) The highest possible output was usually demanded, as represented by the factors given in Fig. 9. (2) The lowest possible sensitivity to changes in the friction value of the linings was required, and this was shown as the factor a t 0 . 4 ~ divided by the factor at 0.3~. (3) Equal wear of the two linings was always desirable, and tHe proportional rate of wear was given. It was wise to judge the ultimate stability of a brake in terms of the coefficient of friction at which it spragged, or became self- Equal Equal 3 t 0 1 Equal 3 to 1 at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from AUTHOR\u2019S REPLY ON INTERNAL EXPANDING SHOE BRAKES FOR ROAD VEHICLES 65 applying, and that was determined by the design of the shoe, and not by the brake layout"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.18-1.png",
        "caption": "FIGURE 8.18",
        "texts": [
            " The analysis results were validated using experimental data, acquired by mounting a data acquisition system on the racecar and driving the car on the test track following specific driving scenarios that were consistent with those of the simulations. The results were used to aid the suspension design for handling and cornering (Wheeler, 2006). Assembling an entire vehicle suspension for motion analysis is nontrivial and beyond the scope of this book. Therefore, only the right front quarter of the suspension, as shown in Figure 8.18, is included in this section. The purpose of this case study is mainly to highlight capabilities in Pro/ENGINEER Mechanism Design and SolidWorks Motion for supporting design of the kinematic characteristics of vehicle suspension. The motion model was first created in Pro/ENGINEER for kinematic analysis, and then imported into SolidWorks for dynamic analysis and design studies. The road profile is characterized by the geometric shape of a profile cam, which is assembled to the tire using a cam-follower connection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure6.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure6.10-1.png",
        "caption": "Fig. 6.10 Plate problem with two edges fixed",
        "texts": [
            "135) 310 6 Classical Plate Elements Derive the 12 interpolation functions Ni(x, y) for this element in Cartesian coordinates. 6.7 Second-order Derivatives of Interpolation Functions in Cartesian Coordinates Consider the interpolation functionsNi(x, y) for a classical plate element in Cartesian coordinates, see supplementary problem6.6. Calculate the second-order derivatives in Cartesian coordinates, i.e. the B-matrix. 6.8 Two-element Example of a Plate Fixed at Two Edges Given is a classical plate structure which is fixed at two sides, see Fig. 6.10. The side lengths of the entire structure are equal to 4a \u00d7 2b. The plate is loaded by two single forces 1 2F acting in the middle of the plate structure. The material is described based on the engineering constants Young\u2019s modulus E and Poisson\u2019s ratio \u03bd. Use two classical plate elements (each 2a \u00d7 2b \u00d7 h) in the following to model the problem and to calculate the nodal unknowns in the middle of the plate structure. Simplify the results for the special case \u03bd \u2192 0 and compare these results with the Euler\u2013Bernoulli solution"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure16-1.png",
        "caption": "Figure 16 Locking mechanism, (a) lock of joint (b) resolution of locking mechanism (c) pistons of locking",
        "texts": [
            " This reduces the friction, and the duplex manipulator can escape from its deadlock. By this mechanism, we can utilise both pushing and pulling operations. The pushing operation is effective for the outside manipulator, and the pulling operation is effective for the inside manipulator. The mechanisms for turning the head link are similar to those of our previous manipulator described in Section 3. Details on the developed manipulator are presented in Section 6. Figure 15 shows the developed duplex manipulator. Figure 16 shows the locking mechanism. Figure 17 shows the decoupled manipulators. Figure 18 shows the pulling mechanism. Table 1 shows the parameters of developed robot. As shown in Figure 16(a), when the hose expands, it pushes the pin until the pin engages with the hole, which locks the joint. Figure 16(b) shows the resolution of the locking mechanism, and Figure 16(c) shows the pistons to expand the hose. As shown in Figure 18, pulling the wire of the locked manipulator pulls the movable manipulator. Figure 19 shows the camera and four LED lights. We employed a UCAM-DLK130T (Elecom) camera, which is used by the operator to search for survivors. We conducted some experiments to demonstrate the effectiveness of the developed duplex manipulator. First, we confirmed the movement of the head link. Figure 20 shows the experimental result for one manipulator. The head could move up to a maximum of about 30 cm in the vertical direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.31-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.31-1.png",
        "caption": "Figure 9.31 The components of Equations (9.80) and (9.81).",
        "texts": [
            " As a function of theus;est is;measRss;est integration time \u0394t, discrete flux linkage can be expressed in terms of xy components. k \u0394t usx k k k 1 (9.76)Rsisx\u03c8 sx est \u03c8 sx est \u0394t usy k Rsisy k (9.77)\u03c8 sy est k \u03c8 sy est k 1 Integration is improved by comparing the stator current estimates isx est and isy est with the measured currents isx and isy. \u0394isx isx isx est (9.78) \u0394isy isy isy est (9.79) Stator flux linkage is determined in terms of rotor angle \u03b8r as follows. \u03c8PMcos \u03b8r (9.80)\u03c8 sx Lsisx \u03c8PMsin \u03b8r (9.81)\u03c8 sy Lsisy This division into components is illustrated in Figure 9.31. Next, stator current estimates can be computed from the rotor angle estimate \u03b8r est k . 1 isx est \u03c8 sx est \u03c8PMcos \u03b8r est k (9.82) Ls 1 isy est \u03c8 sy est \u03c8PMsin \u03b8r est k (9.83) Ls The direct and quadrature inductances are assumed to be equal Ld= Lq= Ls. If inductance is assumed constant, flux linkage becomes a function of current and rotor angle. Therefore, a total differential for the flux linkage can be written. @\u03c8 @\u03c8 \u0394\u03c8 \u0394i \u0394\u03b8r (9.84)s @i @\u03b8r The rotor angle estimate \u03b8r est is corrected by an amount equal to \u0394\u03b8r est until the flux-linkage error \u0394\u03c8 s becomes zero, and \u0394\u03b8r est can be calculated as follows"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000842_amm.665.593-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000842_amm.665.593-Figure1-1.png",
        "caption": "Fig. 1 The rod section in a gas stream",
        "texts": [
            " At study of stability of the drill rods considered in the form of shells, special attention is paid to circular cylindrical shells. They not only meet the requirements of the least weight of the construction but are also prime in manufacture. The purpose of the present paper is examination of stability of the compressed and twisted drill rod presented in the form of a circular cylindrical shell in a supersonic stream of gas. Let us consider the circular cylindrical shell, which is circulated from the outer side by a supersonic stream of gas, subject to strains of torsion and compressing (Fig. 1). The nonlinear equation is examined of not axisymmetric strain of the cylindrical shell of an aspect: ( ) 324 2 2 2 2 2 2 2 2 2 3 2 4 2 2 2 2 1 2 0, 48 hD E w w w w w w w P S A A h x y g x xR x x t P \u03ba\u03b3 \u03ba \u221e +\u2202 \u2202 \u2202 \u2202 \u2202 \u2202  \u2206 \u2206 + + \u2206 + \u2206 + \u2206 + \u2206 + =  \u2202 \u2202 \u2202 \u2202\u2202 \u2202 \u2202   (1) All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-21/05/15,09:12:49) where ( ) 3 212 1 Eh D \u00b5 = \u2212 is the cylindrical rigidity, h width of the shell wall, E Young\u2019s modulus, R radius of the cylindrical shell, 2\u2206 biharmonic operator, w transversal buckling of the shell, \u00b5 Poisson coefficient, P uniformly distributed compressing load, 22 M S R h\u03c0 = tangent loading from the twisting moment M, g\u03b3 \u03c1= \u22c5 specific weight of the material, \u03c1 material density, g free fall acceleration, P\u221e pressure of unperturbed gas, M Mach number, 2M P A h \u03ba \u221e= reduced Mach number, \u03ba polytropic exponent"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003657_s38311-015-0030-0-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003657_s38311-015-0030-0-Figure3-1.png",
        "caption": "FIGURE 3 Electrodynamic actuator",
        "texts": [
            " The electrodynamic actuator generates an output inertial force which is passed through the vehicle structure. The input sensor (either microphone or accelerometer) measures vehicle responses, either in g or dB. The actuator is the system component that generates the counter vibrations. It has to be small enough to fit yet strong enough to provide sufficient force output over the entire operating frequency range. The electrodynamic actuator chosen for this investigation consists of a mass suspended elastically between two steel springs, FIGURE 3. The mass contains a strong permanent magnet made of neodymium or samarium cobalt material. This mass is moved back and forth through the electromagnetic field generated by a surrounding coil. The coil system, the moving mass, the magnet type, the spring path and the springs are all set up in accordance with the frequency response required and the force required over this frequency. Vibrations in cars typically occur within a frequency range of 15 to 480 Hz and with a peak force value of 50 to 200 N"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.36-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.36-1.png",
        "caption": "FIGURE 6.36",
        "texts": [
            " During an initial static analysis of the full vehicle, to settle at kerb height, the rack moves down with the vehicle body relative to the suspension system. This has a pulling effect, or pushing according to the rack position, on the tie rod that causes the front wheels to steer during the initial static analysis. The solution to this is to establish the relationship between the steering column rotation and the steer change in the front wheels and to model this as a direct ratio using two coupler statements to link the rotation between the steering column and each of the front wheel joints as shown in Figure 6.36. Advanced steer-by-wire research platforms, such as the \u2018P1\u2019 vehicle in use at Stanford University, use this layout in a physical vehicle with a software controller solving the constraint equation by driving electric motors to move the steered road wheels. Toe change in front wheels at static equilibrium for simple models. Coupled steering system model. In order to implement the ratios used in the couplers shown in Figure 6.36, linking the rotation of the steering column with the steer change at the road wheels, it is necessary to know the steering ratio. At the start of a vehicle dynamics study the steering ratio can be a model design parameter. In the examples here a ratio of 20 of handwheel rotation to 1 of road wheel steer is used. On some vehicles this may be lower and on trucks or commercial vehicles it may be higher. To treat steering ratio as linear is a simplification of the situation on a modern vehicle. For example the steering ratio may vary between a lower value on centre to a higher value towards the limits of rack travel"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001453_icems.2014.7013831-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001453_icems.2014.7013831-Figure6-1.png",
        "caption": "Fig. 6 Voltage vectors applied in the second process.",
        "texts": [
            " For each voltage vector, the corresponding d-axis current is measured. Comparing all the measured currents, we can get Id_max1 and In the second process, three new voltage vectors whose angles are m1 - , m1 and m1 + are applied to the motor. For each voltage vector, the corresponding d-axis current is still measured to obtain Id_max2 and m2. Then repeat the second process to get a new Id_max2 until voltage vector 21 is applied to the motor. In this process, the initial value of is 7.5, and for every new round, is halved, as shown in Fig. 6. The flow chart of this method is shown in Fig. 7. To avoid influence on current measuring, all the gate signals of the inverter are turned off for a period of Toff between two adjacent injections to ensure each phase current attenuates to zero. In addition, during every injection, the d-axis current is only sampled in the last several milliseconds instead of the whole time period. As Fig. 8 shows, most estimation results are close to actual initial rotor positions, except two estimated angles which are 180\u00b0 ahead of their actual position"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001174_0954405414554016-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001174_0954405414554016-Figure7-1.png",
        "caption": "Figure 7. Geometry of the rolling body positions: (a) relationship between rolling and inner and outer ring, (b) low speed and light load (axis preloading), and (c) low speed and heavy load (axis and radial preloading).",
        "texts": [
            " Ri1 = Db 2 , R0i1 = fiDb Ri2 = Db 2 , R0i2 = Di +Do 2Db cos a0 4 cos a0 ai = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ri1R 0 i1 Ri1 R0i1 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Rim 3 r bi = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ri2R 0 i2 Ri2 +R0i2 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Rin 3 r Ri = ffiffiffiffiffiffiffiffiffiffiffiffiffiffi RimRin p aibi =Ridi Ki = 4 3 ER 1=2 i d 1=2 i = 16EF2R2 i 9 1=3 or Kic = 4 3 ER 1=2 i \u00f021\u00de Ro1 = Db 2 , R0o1 = foDb Ro2 = Db 2 , R0o2 = Di +Do +2Db cos a0 4 cos a0 ao = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ro1R 0 o1 R0o1 Ro1 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Rom 3 r bo = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ro2R 0 o2 R0o2 Ro2 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ron 3 r Ro = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi RomRon p aobo =Rodo Ko = 4 3 ER1=2 o d1=2 o = 16EF2R2 o 9 1=3 or Koc = 4 3 ER1=2 o \u00f022\u00de at NANYANG TECH UNIV LIBRARY on June 5, 2016pib.sagepub.comDownloaded from Integrated contact stiffness k= 1 1=Ki +1=Ko or kc = 1 1=Kic\u00f0 \u00de2=3 + 1=Koc\u00f0 \u00de2=3 !3=2 \u00f023\u00de Axial and radial stiffnesses of the bearing. Figure 7 shows specific aspects of the contact geometry for a rolling body and those changes in its position relative to the inner and outer rings for different states of the system under study. Geometrical and physical relationships A= fi + fo 1\u00f0 \u00deDb sin a= A sin a0 + da A sin a0 + da\u00f0 \u00de2 + A cos a0 + dr cos c\u00f0 \u00de2 h i1=2 cos a= A cos a0 + dr cos c A sin a0 + da\u00f0 \u00de2 + A cos a0 + dr cos c\u00f0 \u00de2 h i1=2 Fa = XNb j=1 Qcj sin a Fr = XNb j=1 Qcj cos cj cos a \u00f024\u00de Under low speed and light load d A 1+ d2 a +2 da sin a0\u00f0 \u00de 1=2 1h i =0 sin a\u00f0 \u00de 1+ d2 a +2 da sin a0\u00f0 \u00de 1=2 sin a0\u00f0 \u00de+ da =0 Fa=Nb kcd 3=2 sin (a)=0 \u00f025\u00de Under low speed and heavy load at NANYANG TECH UNIV LIBRARY on June 5, 2016pib"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.22-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.22-1.png",
        "caption": "Fig. 3.22 Spatial four-bar mechanism",
        "texts": [],
        "surrounding_texts": [
            "For the position analysis of the nonlinear transmission motion, e.g. the rotation of the input crank (angle b1) can be chosen as the independent input coordinate q \u00bc b1. The deflection of the right output lever (angle b7) such as the motion of the connecting coupling bar d can be calculated. In particular, one can show that the output variable, which is to be calculated in the first step, b7 \u00bc b7\u00f0b1; geometry of the initial position\u00de; \u00f03:14\u00de can be found explicitly through the solution of the equation Fig. 3.23 Schematic representation of the spatial four-bar mechanism as transmission element\u2014 kinematic transformer A cos b7 \u00fe B sin b7 \u00fe C \u00bc 0: \u00f03:15\u00de For this equations, there exist, based on three coefficients A, B, and C; which are dependent on b1 and on the geometry of the initial position of the mechanism, normally two real solutions, only one of which fulfills the constraints of the initial position of the kinematic system. The unknown variables in the four-link mechanism, which only need to be calculated if necessary, are the two angles of the CARDAN-joint b2; b3 and the three angles of the spherical joint b4; b5; b6. Assuming these two joints, corresponding to Table 3.2, can be constructed from revolute joints, then these angles can be recursively calculated. For further details, see (Woernle 1988). The overall solution structure of the non-linear position analysis of the spatial four-link mechanism can be represented as follows: g1 b1; b7\u00f0 \u00de \u00bc 0; g2 b1; b7; b2\u00f0 \u00de \u00bc 0; g3 b1; b7; b2; b3\u00f0 \u00de \u00bc 0; g4 b1; b7; b2; b3; b4\u00f0 \u00de \u00bc 0; g5 b1; b7; b2; b3; b4; b5\u00f0 \u00de \u00bc 0; g6 b1; b7; b2; b3; b4; b5; b6\u00f0 \u00de \u00bc 0: 9 >>>>= >>>; \u00f03:16\u00de The distinguishable recursive solution structure of the system of constraint equations in Eq. (3.16) somewhat becomes clearer when one observes the JACOBIAN-Matrix Jb of the constraint equations, which is also the basis for the velocity-analysis: Jb \u00bc og ob \u00bc 0 2 6666664 3 7777775 g1 g2 g3 g4 g5 g6 b7 b2 b3 b4 b5 b6 : \u00f03:17\u00de This JACOBIAN matrix Jb possesses a lower-triangle structure and corresponds to a recursively explicit solution of the position analysis of the four-link mechanism, as shown in Li (1990). The complete kinematic transmission characteristics of the four-link mechanism on position level, as well as on the velocity and acceleration level can be combined in the non-linear transmission element, which represents now the complete version of the already introduced kinematic transformer (Fig. 3.25). For more details on the kinematics of four-link mechanisms see the following references (Hiller 1981; Woernle 1988; Kecskemethy 1993)."
        ]
    },
    {
        "image_filename": "designv11_64_0000497_9781119011804.ch6-Figure6.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000497_9781119011804.ch6-Figure6.1-1.png",
        "caption": "FIGURE 6.1 Ray optics model of optical force on a dielectric microsphere. (a) Illustration of a single incident ray\u2014geometry, optical force, and force components Fgrad and Fscat. (b) Illustration of a focused beam showing total Fgrad and Fscat. (c) Typical plot of optical force versus displacement with a linear region for small displacements. (d) Elastic spring model for small displacements with stiffness kx, ky, kz.",
        "texts": [
            " Dielectric microspheres have been used since Ashkin\u2019s pioneering experiments [3], and they remain as preferred structures for optical trapping research, partly owing to their practical accessibility when performing experiments as well as the tractability of a dielectric microsphere model for theoretical analyses. This model is relevant for cell handling, either as a simplified model of the cell itself or for the actual modeling of trapped dielectric microspheres that are often used as handles for indirectly manipulating cells with light. A simplified ray optics model is shown in Figure 6.1, which illustrates the optical force that arises when light interacts with a dielectric sphere, here manifesting as refraction at the interface. In this picture, the optical force balances the rate of optical momentum change due to light deflections in accordance with the momentum conservation principle. In the ray model (valid for larger objects, size \u226b \ud835\udf06), the optical force may be decomposed into components parallel and orthogonal to the incident ray, which may be denoted as scattering force and gradient force, respectively [9]",
            "1) The quality factor, Q, is a measure of how well the system generates force, which depends on how much it alters the incident light momentum flux, nmP\u2215c (where c\u2215nm is the speed of light in the surrounding material)1. For example, perfect plane-wave reflection at normal incidence corresponds to Q = 2, while total absorption leads to Q = 1. Achieving high quality factors is highly desirable, especially when handling cells, to get the most force, while keeping modest input powers, and to avoid parasitic effects from high intensities. For a beam incident in water (nm = 1.33), a quality factor of Q = 1 can exert 4.44 pN for every 1 mW of incident power. For the trapped microsphere illustrated in Figure 6.1, the optical force and quality factor depend on the bead position (typically similar to the simplified version shown in Figure 6.1c). For small displacements, the optical force can be modeled as an elastic spring obeying Hooke\u2019s law, F = \u2212ks. The optical force decays with further displacements beyond a characteristic displacement comparable to the bead radius2. For comparison, we may take, as a point of reference, Ashkin\u2019s estimates, based on ray optics model, that found \u22120.276 < Qaxial < 0.490 for a dielectric sphere that is axially translated along its equilibrium point (relative refractive index = nsphere\u2215nm = 1.2, e.g., polystyrene in water, NA = 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002424_s1068798x16040109-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002424_s1068798x16040109-Figure3-1.png",
        "caption": "Fig. 3. MarShaft MAN manual instrument (a) and DMS 120 measuring head (b).",
        "texts": [
            " We now consider the instruments produced by Mahr (Germany), which is one of the world\u2019s leading instrument makers. Mahr instruments may be divided into manual devices and automatic systems with a servo drive. The manual instruments are characterized by lower pro ductivity but lower cost. The automatic systems with a servo drive (tactile and optical models) are signifi cantly more productive and do not require operator participation in the measurements (with consequent minimization of human error) but are more expensive. MANUAL INSTRUMENTS The MarShaft MAN instrument has a horizontal configuration (Fig. 3a). There are two zones on a steel or granite frame 3. High precision steel guides are mounted in one zone. Carriages 4, which carry the measuring heads, move over the guides. The front 1 and rear 2 headstocks run on separate guides in the second zone. Rear headstock 2 may move longitudi nally on its guide. Front headstock 1 is positioned on the guides with high precision relative to rear head stock 2. Standards for calibration of the system are mounted on rear headstock 2. This position of the standard significantly simplifies the operation of the system. The precision may be verified at any time and, where necessary, the system may be recalibrated. As a rule, three measuring heads are mounted on carriages 4: heads for measuring the diameter, the length, and the wobble. The heads are such that, for each measurement, the minimum number of actions is required. To measure the diameter using the DMS 120 head (for diameters up to 120 mm), we simply pull the rear part of the head over it (Fig. 3b). The measuring tips move toward the axis of the part until they are stopped at the measured surface. In the extreme rear position, the head creates the required measuring force and the result is recorded. Thus, to measure diameters, a single motion is required. Therefore, such a measurement takes 3\u20135 s, including the positioning of the head. To measure diameters up to 220 mm, we use a DMS 220 head with motorized tip motion. To measure lengths, we use the LTS head. Its oper ating principle is as follows"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure4-1.png",
        "caption": "Figure 4. System of two concurrent mutually orthogonal dampers and the coordinate system used.",
        "texts": [
            " Thus, one needs to define two of them a priori and then calculate the rest of them from equations (17a,b) and (18a,b). Corollary 2.1. The sum of the stiffness coefficients of two types of equivalent systems of springs is given by the following relationship kI \u00fe kII \u00bc k 0 I \u00fe k 0 II \u00fe k0I I lI \u00fe k0II II lII \u00f019\u00de Proof. This follows directly from equation (18a,b). Let us consider now the system of concurrent dampers from Figure 1(a) and define the equivalent system of two mutually orthogonal dampers shown in Figure 4. Theorem 3. The system of N concurrent in-plane linear viscous dampers can be replaced by the equivalent system of two mutually orthogonal linear viscous dampers, where the position of one of them with respect to the horizontal is given by the angle c that satisfies tan 2 c \u00bc PN j\u00bc1 cj sin 2 \u2019jPN j\u00bc1 cj cos 2 \u2019j \u00f020\u00de where cj is the damping coefficient of each of the original dampers and \u2019j is the angle between each of them and the horizontal. The damping coefficients of the new dampers are cI \u00bc XN j\u00bc1 cj cos 2 \u2019j c , cII \u00bc XN j\u00bc1 cj sin 2 \u2019j c \u00f021a; b\u00de Proof"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001180_icra.2014.6906999-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001180_icra.2014.6906999-Figure2-1.png",
        "caption": "Figure 2. A speed graph edge in the vicinity of a polygonal obstacle.",
        "texts": [
            " In order to find the true time optimal path between two speed graph nodes, a saturation method is proposed and explained in the technical report [13]. The speed graph in next extended to polygonal environments. Each polygonal obstacle perimeter consists of line segments joined by vertices. These can be treated as individual point or wall segment obstacles called obstacle-features. When the robot travels between polygonal obstacles, it is in the vicinity of a single obstacle-feature ( a vertex or edge) at each instant along its path. Consider two nodes that lie on the boundary of a common Voronoi cell (e.g. q12 and q13 in Figure 2). To construct the time optimal arc between these nodes, consider the proximal obstacle-features of the polygonal obstacle at the center of the Voronoi cell (e.g. O1 in Figure 2). Assume the speed graph edge crosses l equidistant lines spanned between adjacent obstaclefeatures of the polygonal obstacle. The speed graph edge is consequently affected by l+1 feature-obstacles (e.g. l+1=3 obstacle-features in Figure 2). The speed graph edge can thus be parameterized by the points p1, ..., pl where the edge crosses the equidistant lines. Each crossing point pi can freely vary along the i\u2019th equidistant line, but only in the segment bounded by the obstacle and the Voronoi edge. Since each speed graph edge is required to be time optimal, it must possess a continuous tangent at p1, ..., pl . Computing the optimizing values of p1, ..., pn is done by solving a suitable convex optimization problem (see [13]). This section describes a scheme for computing the time optimal path within a specific homotopy class suggested by the speed graph"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.5-1.png",
        "caption": "Fig. 10.5 B\u2013H characteristics: (a) of free space or air, (b) of a typical magnetic core material",
        "texts": [
            " So for the example of Fig. 10.4, Eq. (10.7) reduces to F (t) = H(t) m = i(t) (10.8) Thus, the magnetic field strength H(t) is related to the winding current i(t). We can view winding currents as sources of MMF. Equation (10.8) states that the MMF around the core, F (t) = H(t) m, is equal to the winding current MMF i(t). The total MMF around the closed loop, accounting for both MMFs, is zero. The relationship between B and H, or equivalently between \u03a6 and F , is determined by the core material characteristics. Figure 10.5a illustrates the characteristics of free space, or air: B = \u03bc0H (10.9) The quantity \u03bc0 is the permeability of free space, and is equal to 4\u03c0 \u00b7 10\u22127 Henries per meter in MKS units. Figure 10.5b illustrates the B\u2013H characteristic of a typical iron alloy under highlevel sinusoidal steady-state excitation. The characteristic is highly nonlinear, and exhibits both hysteresis and saturation. The exact shape of the characteristic is dependent on the excitation, and is difficult to predict for arbitrary waveforms. For purposes of analysis, the core material characteristic of Fig. 10.5b is usually modeled by the linear or piecewise-linear characteristics of Fig. 10.6. In Fig. 10.6a, hysteresis and saturation are ignored. The B\u2013H characteristic is then given by B = \u03bcH \u03bc = \u03bcr\u03bc0 (10.10) The core material permeability \u03bc can be expressed as the product of the relative permeability \u03bcr and of \u03bc0. Typical values of \u03bcr lie in the range 103 to 105. The piecewise-linear model of Fig. 10.6b accounts for saturation but not hysteresis. The core material saturates when the magnitude of the flux density B exceeds the saturation flux density Bsat"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001841_978-3-319-21266-1_26-Figure26.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001841_978-3-319-21266-1_26-Figure26.1-1.png",
        "caption": "Fig. 26.1 System overview of the environment employed in this work. a Overview of the transfer station showing the actuators of our setup. b Detailed view of the setup where the robot currently unloads an AGV",
        "texts": [
            " 2-4, 44227 Dortmund, Germany \u00a9 Springer International Publishing Switzerland 2016 U. Clausen et al. (eds.), Commercial Transport, Lecture Notes in Logistics, DOI 10.1007/978-3-319-21266-1_26 397 Additionally to the manually operated picking stations, the system exhibits an industrial robot which is able to pick or place cuboid shaped goods from or to moving AGVs. The robot\u2013equipped with a vacuum gripper and a tool-mounted 3D camera\u2013is located in the center of a transfer station bordered by a safety fence (cf. Fig. 26.1a). AGVs are carrying bins, entering, and leaving the transfer station area on freely programmable routes. Additionally, the robot cell contains a stationary 2D camera and a (de-) palletizing area. Within this setting, the robot is picking a packet from a moving AGV (cf. Fig. 26.1b). Hence, two major problems have to be approached: First, the moving AGV and its loading needs to be detected and localized. Second, the detected loading needs to be grasped on-the-fly. The latter includes following the AGV with the robot arm by predicting its position in the near future, grasping the packet at a reasonable point in time during the tracking, and depositing the packet at the unloading area (see Fig. 26.1a). The process of unloading is related to the bin-picking problem albeit with moving charge carriers. Bin-picking in static scenarios is a well-known problem in research\u2014several approaches and solutions came up in the late 80s and early 90s (cf. Al-Hujazi and Sood 1990; Bolles and Horaud 1986). In the meantime, many mobile robot platforms have been developed. Equipped with manipulators and camera systems, they are able to grasp objects out of non-moving boxes (Nieuwenhuisen et al. 2013). Thus, the moving part here is the robot, not the bin",
            " We assume that the AGV carries a cuboid shaped boxboard having a color not equal to the color of the AGV which seems reasonable since most packagings have this sort of appearance. Figure 26.7 visualizes the processing steps of a point cloud P R 6 retrieved by the 3D camera. It captures point clouds (see Fig. 26.8a) from the topside of the AGV which get preprocessed by dynamic distance filtering to remove the ground points pi \u00bc xi; yi; zi; ri; gi; bi\u00f0 \u00deT2 P, voxelgrid downsampling to reduce the size of the input while still preserving the geometric structure of the AGV, color filtering to remove the salient colored body elements of the AGV (orange in Fig. 26.1), and statistical outlier removal to suppress scattered points in the residual point set. By detecting planes in the point cloud, we obtain an approximation of the packet transported by the AGV (cf. Sect. 26.3.1). Finally, the orientation of the packet is determined to properly align the vacuum gripper (see Sect. 26.3.2) for grasping. Again, like in Sect. 26.2, positions need to be predicted in the absence of point cloud data which occurs when the robot actually grabs the loading. Details are explained in Sect",
            " The predicted position xM\u00f0t\u00de is used to grasp the packet. This section deals with the experimental results of the presented concepts. First, the system setup is presented in detail. Second, the accuracy of the 2D AGV detection is elaborated (see Sect. 26.4.1). Third, the accuracy of the 3D load detection is analyzed (see Sect. 26.4.2). Finally, the overall system performance is presented involving all algorithms described in this paper (see Sect. 26.4.3). As already introduced in Sect. 26.1, the transfer station (cf. Fig. 26.1) is essentially composed of two types of actuators: an industrial robot of type KUKA KR 125/3 equipped with a vacuum gripper tool (SCHMALZ FX-400/12C-SV) and the AGVs of the Multishuttle Move system (Kamagaew et al. 2011). They store and transport a single boxboard with a cuboid shape and a solid color. The sensors in this setup are the (stationary) 2D CCD camera DFK 31BU03. H produced by The Imaging Source GmbH with a resolution of 1024 \u00d7 768 px at 30 fps and the ASUS Xtion Pro Live. The latter provides a resolution of 640 \u00d7 480 px at 30 fps and a working range of 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001380_sami.2015.7061901-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001380_sami.2015.7061901-Figure3-1.png",
        "caption": "Figure 3 Test plate dimensions",
        "texts": [
            " Characteristics of the test plates The concept \u201ccharacteristics of the test plate\u201d can be understood as the determining/defining of the dimensional circumference of the given test plate; however, the dimensional circumferences are to a large extent limited by the working space of the equipment for applying the layers, or the device for coating the test plates. The dimensional circumferences are also limited by the anticipated implantation itself (its specific type) with regard to determining the measure of compatibility, whether with the use of the same technological coating or without it. On the basis of the above-mentioned limiting factors, the test plates were modelled in a 3D CAD modelling software called \u201cSolidWorks 2012\u201d, whereby the given test plates typically have the dimensions: length = 10.0 mm, width = 0.50 mm and height = 5.0 mm (Fig. 3). B. Characteristics of the building platform The characteristics of the platform, or its dimensions and the material from which it is made, are exactly determined by the maker and seller of the specific device itself, or a machine using the DMLS technology for creating the product on the basis of the determined requirements and needs (Fig.4). C. Method of orientating the test plates on a platform After considering the conditions or rules for correct orientation of the model/product on the platform, several methods of orientation of the test plates were proposed (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003244_s1068366616040073-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003244_s1068366616040073-Figure5-1.png",
        "caption": "Fig. 5. Plotting slip-line field using numerical method (D = 106 mm, \u03b4 = 1.05, \u03c61 = 21.9\u00b0, \u03c62 = 7.9\u00b0, and \u0394\u03c6 = 20.46\u00b0).",
        "texts": [
            ", an increase in the slippage expressed in the relative parameters (vX/vZ) leads to an increase in the force of friction. As a result, we can consider it to have been proved that the self-setting of the forces of friction during CR is governed by the dimensionless parameter, i.e., the ratio of the velocity of the tool to the velocity of the flow of the metal in the near-contact layer of the deformation zone. Let us consider the relationship between the velocities near the contact zone and Prandtl\u2019s coefficient of friction \u03bc\u03c4, which was obtained using the slip-line technique (Fig. 5). We carry out an analysis of the dependence of Prandtl\u2019s coefficient of friction \u03bc\u03c4 on the change in the relative velocity of the f low of the metal vtool/vcont (Fig. 6). The velocity of the tool vtool remains constant, while the velocity of the f low of the metal in the near-contact layer vcont decreases when 328 JOURNAL OF FRICTION AND WEAR Vol. 37 No. 4 2016 KOZHEVNIKOVA passing from point K to point N; subsequently, this velocity increases when passing from point N to point \u041c, decreases at point \u041c because of the primary effect of the velocity discontinuity v3, then remains unchanged until it achieves point L (Fig. 5). Since the relative parameter vtool/vcont is the reciprocal of the velocity of the f low of the metal, the dependence of this parameter on the length of the contact zone is inverse to the dependence of vcont. In this case, the direct dependence of the coefficient of friction \u03bc\u03c4 on the parameter vtool/vcont is observed; i.e., the greater the slippage, the higher the coefficient of friction. Thus, the results of the experiments and the theoretical study, which has been carried out using the upper limit estimate method and the slip-line technique, make it possible to unambiguously identify the mechanism of the self-setting of the forces of friction during CR via the self-governing of the slippage of the metal being deformed against the tool"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.61-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.61-1.png",
        "caption": "Fig. 3.61 Squared frame structure in a flat and b angled orientation. Top full model; middle half model and bottom quarter model",
        "texts": [
            " \u2022 Simplify your general solution for the special case that the beam is absent. 182 3 Euler\u2013Bernoulli Beams and Frames Derive the stiffnessmatrixKe XY (\u03b11,\u03b11) for a generalized beam (Bernoulli) element which can deform in the global X\u2013Y plane. Consider that the rotation angle between the local and global coordinate system is different at both nodes. 3.37 Mechanical properties of squared frame structure Given is a squared frame structure made of generalized beam elements with side length L as shown in Fig. 3.61. Two different orientations, i.e. flat or angled orientation, should be considered in the following. Calculate the displacement (BC: F) or reaction force (BC: u) of the point of load application and estimate the macroscopic stiffness Estruct of the frame structure. Simplify your results for the macroscopic stiffness for the special case A = \u03c0d2 4 and I = \u03c0d4 64 , i.e. a circular cross section of the beam elements. The derivation should be performed first for the full model, then for the half model and finally for the quarter model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.9-1.png",
        "caption": "FIGURE 8.9",
        "texts": [
            " The kinematic joints allow motion in some directions and constrain it in others. The types of motion allowed and constrained are related to the characteristics and intended use of the joint, which can be usually characterized by the degrees of freedom it allows. For a planar mechanism, there are two kinds of joints: planar J1 joints that allow one DOF (restrict two DOF); and planar J2 joints that allow two DOF (restrict one DOF). The revolute and slider joints discussed before are J1 joints and the pin-in-slot is a J2 joint, as shown in Figure 8.9. Threedimensional or spatial joints are classified into two categories based on the type of contact between the two members making a joint: lower pair joint and higher pair joint. The contact can be point, line, or area. A third category of kinematic joint comprises the joints formed by combining two or more lower pair and/or higher pair joints. Such joints are termed compound joints. The two members forming a lower pair joint have area contact between the two mating surfaces. The contact stress is thus smaller for lower pair joints as compared to higher pair joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003196_cistem.2014.7076970-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003196_cistem.2014.7076970-Figure11-1.png",
        "caption": "Fig. 11 .Distribution du champ magn\u00e9tique dans le plan (\u03b8, Z) \u00e0",
        "texts": [
            " La simulation a \u00e9t\u00e9 r\u00e9p\u00e9t\u00e9e pour d\u00e9terminer quelle longueur de culasse statorique nous donne l'induction radiale maximale. Puis la structures r\u00e9duite \u00e0 \u00e9t\u00e9 compar\u00e9e \u00e0 la structure initiale de r\u00e9f\u00e9rence ou l'induit occupe toute la longueur entre les deux sol\u00e9no\u00efdes. La Fig. 10 repr\u00e9sente deux courbes de l'induction magn\u00e9tique radiale Br (Z). Apr\u00e8s avoir calcul\u00e9 la valeur moyenne de l'induction radiale utile, pour chaque configuration, on obtient une induction de 1,72 T dans la topologie A et 2,06 T pour B (Fig. 10). La Fig. 11 repr\u00e9sente la r\u00e9partition de la composante radiale du champ magn\u00e9tique sur une surface dans le plan (\u03b8, z), dans la zone utile de la machine en r= Rext+2cm. Cette distribution pr\u00e9sente les deux p\u00f4les nord et sud obtenus \u00e0 partir de l'inducteur \u00e9tudi\u00e9. IV. EXP\u00c9RIMENTATION Nous avons d\u00e9cid\u00e9 de r\u00e9aliser une machine bas\u00e9e sur ce principe. Pour des raisons de co\u00fbt les bobines supraconductrices seront r\u00e9alis\u00e9es en NbTi. Le NbTi est un mat\u00e9riau supraconducteur \u00e0 basse temp\u00e9rature critique refroidi \u00e0 l'h\u00e9lium liquide \u00e0 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002249_ijmmm.2015.073153-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002249_ijmmm.2015.073153-Figure5-1.png",
        "caption": "Figure 5 Geometric parameters of pyramids when in use",
        "texts": [
            " This relation defines the speed of indentation by the abrasive tool in the polished material (dz/dt). This speed is the product of a constant (Cp) specific to the polished material, the nature and size of the grains, the characteristics of the binder, the pressure (p) exerted by the abrasive grains on the polished surface and the feed speed of the grains (V). p dz C p V dt = \u22c5 \u22c5 (4) Taking into account the relative movement of the abrasive grains as they sweep a surface (Sprojected), if pressure (p) is exerted at any point on the apex of the pyramid (Scontact) by the abrasive grains [see Figure 5(a)], then the larger the abrasive-material contact area, the more material will be removed. Preston\u2019s model as applied to the pyramid-shaped abrasive belts is given in equation (5). contact p projected dz S C p V dt S = \u22c5 \u22c5 \u22c5 (5) Considering the geometry of the pyramids when in use, as shown in Figure 5(b), the surface ratio is expressed in equation (6). 2 2 2(1 ) ,contact projected S x h z k S d h \u2212\u239b \u239e \u239b \u239e= = = \u2212\u239c \u239f \u239c \u239f \u239d \u23a0 \u239d \u23a0 (6) where k represents the rate of wear of the pyramids. The pyramids are new when k = 1, and fully worn when k = 0. Equation (5) then becomes: 2(1 ) p dz k C p V dt = \u2212 \u22c5 \u22c5 \u22c5 (7) From this relation, the material removal rate can be expressed: projected projected p dv dzS S C p V dt dt = \u22c5 = \u22c5 \u22c5 \u22c5 (8) Pressure (p) derives from the thrust force (Fz) applied to a pad with a constant cross-section pressing the abrasive grains downwards onto the material to be polished"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure1-1.png",
        "caption": "Fig. 1. 3D model of the planned design",
        "texts": [],
        "surrounding_texts": [
            "Index Terms \u2013 conductor car, innovation, live-line maintenance, high voltage\nI. INTRODUCTION\nIn the Hungarian high voltage grid there is an increasing need to repair and change composite insulators while a power line is energized. In the High Voltage Laboratory of Budapest University of Technology and Economics a new type of conductor car is under development. There are some technical and economic aspects which are promised to be more advantageous than in case of the current designs. One of the most significant difference will be the notably decreased weight of the equipment. Another important criterion that the new type of conductor car has to be capable to pass through the suspension insulators. New design, new kinds of applied materials make the manufacturing more economical. In case of successful operational, mechanical and electrical tests manufacturing and trading may also begin shortly. With a co-operation of the Hungarian transmission system operator a demonstration of the capabilities of the new equipment is also planned in the Hungarian grid. With the use of this new type of conductor car, high voltage live-line maintenance may become more safe, efficient and economic.\nII. MOTIVATION OF DEVELOPMENT\nSimilarly to the international trends [1], role of high voltage live-line maintenance is increasing continuously. In Hungary, there is a need from TSO side to develop an equipment for executing advanced works (such as composite insulator change) or to pass through suspension insulators. Many of the maintenance works at high voltage levels are required to be executed with live-line techniques. Currently a part of these activities often cannot been performed because of the lack of required equipment and/or technology. As a part of an international project in the High Voltage Laboratory of Budapest University of Technology and Economics, a new type of conductor car is under development to make more and more maintenance works possible while the grid is energized. This new type of equipment is planned to be capable to perform new kinds of tasks above the currently executable live-line activities with the existing conventional designs.\nIII. ADVANTAGES OF NEW DESIGN AND MATERIALS\nAPPLIED\nAs the result of the new kind of applied materials, conductor car will be more economical to produce and will be lighter then currently available models on the market. Not only the material itself, but the principle of design is new: while conventional conductor cars are made of conductive material, this one consists insulated parts only. Although it does not contain any conductive parts, it must have been taken into account as a conductive object because of the non-qualified insulation for high voltage and the presence of the worker, who is supposed to be completely conductive. These properties require a whole new way of approach during planning, design, manufacturing and operation as well.\n978-1-5090-5165-6/16/$31.00 \u00a92016 IEEE",
            "IV. MECHANICAL SIMULATIONS\nFor the mechanical simulations finite element solution of ANSYS has been applied. The main goal of the simulation was to determine the mechanical safety factor of the new design. For this, forces and torques were determined in each node of a 3D CAD model representing the proposed design. Fig. 2 shows the 3D model created for the mechanical simulations. In Fig. 3 identification of each node is marked.\nBoth forces and torques were analyzed in case of different mechanical loading conditions. In the first two cases the mechanical load was supposed to be symmetric in the middle of the conductor car (1500 N). In the first case the elastic modulus of the applied material was chosen as 5 GPa, while in the second case it was 10 GPa. In the third case the loading conditions were asymmetric, all the mechanical loads were concentrated to the corner of the conductor car, while the elastic modulus of the material was 5 GPa. Deformation of the structure in each cases is shown in Fig. 4-6.\n978-1-5090-5165-6/16/$31.00 \u00a92016 IEEE",
            "One of the most important questions regarding to the mechanical loading conditions of any structure is the distribution and the maximal value of forces affecting to the different nodes. Fig. 8 shows the distribution of the forces in the model.\nThe maximal force value was 611.48 N(y) in case #1, 754.86 N(y) in case #2 and 1718.55 N(y) in case #3. It can be determined that asymmetry has a notable effect on the magnitude of maximal forces affecting to the structure. In case of symmetric loading conditions, none of the partial forces reached value of the total mechanical load (1500 N).\nAnother important aspect regarding to the mechanical conditions of the new design is the distribution of the torques. Fig. 9 shows the magnitudes for each cases in directions x, y and z.\nIn case #1 the maximal value of torque was 10,99 Nm(y), while in case #2 and case #3 it was 11,82 Nm(y) and 26,8 Nm(y), respectively. As it can be seen similarly to the forces, symmetrical loading is very important regarding to reach nearly equal stresses. Maximal asymmetric load is 15% higher than nominal value. To determine the mechanical safety factor of the new design, 25 connecting element samples were inspected by the Department of Polymer Engineering of Budapest University of Technology and Economics\u2019 Faculty of Mechanical Engineering.\n978-1-5090-5165-6/16/$31.00 \u00a92016 IEEE"
        ]
    },
    {
        "image_filename": "designv11_64_0003725_978-3-658-16176-7-Figure95-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003725_978-3-658-16176-7-Figure95-1.png",
        "caption": "Figure 95 - Markov chain",
        "texts": [
            " = { , \u2026 , } (10-174) The observer can only observe the generated sequence of symbols = , \u2026 , (10-175) where \u2208 (10-176) The states that have led to the output of the sequence of symbols are unknown, or hidden, thus such a model is called hidden Markov model. The output of the symbols is arbitrary and only depends from the state , never from formerly taken states and output symbols: ( | \u2026 , \u2026 ) = ( | ) (10-177) In each state each Symbol may be put out, with the probabilities = (10-178) where 320 10 Reverse engineering the mind = | (10-179) Furthermore, (\u2200 , ) \u2265 0 (10-180) and (\u2200 ) \u2211 = 1 (10-181) The hidden Markov model is fully defined by the parameters ( , , ). Figure 95 - Markov chain and Figure 96 - Hidden Markov model graphically describe the differences between the two approaches - describe the states. The algorithm for the generation of the sequence of symbols = , \u2026 , is as follows: Start 1. Set = 1 and select an initial state = under consideration of . 2. Repeat a) Select an observation symbol = under consideration of ( | ) from the matrix B. b) If < pass over to the state = under consideration of the matrix . else End process. c) Set = + 1 3. Until has been reached"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002380_transjsme.15-00563-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002380_transjsme.15-00563-Figure5-1.png",
        "caption": "Fig. 5 Propeller rotation at Yaw turn motion",
        "texts": [],
        "surrounding_texts": [
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u306f\u3053\u308c\u306b\u52a0\u3048\uff0c\u8155\u90e8\u306e\u6298\u308a\u305f\u305f\u307f\u65b9\u5411\u3068\u30d2\u30f3\u30b8\u69cb\u9020\u306b\u7740\u76ee\u3059\u308b\u3053\u3068\u3067\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306b\u3088\u308b\u81ea\u5df1\u5c55\u958b\u304c\u53ef\u80fd\u306a\u69cb \u9020\u3092\u63d0\u6848\u3059\u308b\uff0e\u63d0\u6848\u3059\u308b\u69cb\u9020\u3067\u306f\uff0c\u5c55\u958b\u30fb\u56fa\u5b9a\u306e\u305f\u3081\u306e\u98db\u884c\u306b\u76f4\u63a5\u95a2\u4fc2\u7121\u3044\u30a2\u30af\u30c1\u30e5\u30a8\u30fc\u30bf\u304c\u4e0d\u8981\u3068\u306a\u308a\uff0c\u6a5f\u4f53\u91cd \u91cf\u3092\u6291\u3048\u3089\u308c\u308b\uff0e\n\u63d0\u6848\u306b\u57fa\u304d\u8a2d\u8a08\u3057\u305f\u30d2\u30f3\u30b8\u90e8\u306e\u8a73\u7d30\u3092\u56f3 3\u306b\u793a\u3059\uff0e\u3053\u306e\u30d2\u30f3\u30b8\u306f\u30b9\u30e0\u30fc\u30ba\u306b\u53d7\u52d5\u52d5\u4f5c\u3059\u308b\u305f\u3081\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u304c \u52d5\u4f5c\u3057\u3066\u3044\u306a\u3044\u5834\u5408\uff0c\u91cd\u529b\u306b\u3088\u308a\uff0c\u8155\u90e8\u306f\u4e0b\u306b\u6298\u308c\u66f2\u304c\u308b\uff08\u56f3 3-\u5de6\uff09\uff0e\u4e00\u65b9\uff0c\u30d7\u30ed\u30da\u30e9\u306e\u63a8\u529b\u65b9\u5411\u3068\uff0c\u8155\u90e8\u306e\u5c55\u958b \u65b9\u5411\u3092\u4e00\u81f4\u3055\u305b\u308b\u3053\u3068\u306b\u3088\u308a\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u3092\u56de\u8ee2\u3055\u305b\u3066\u63a8\u529b\u3092\u5f97\u308b\u3053\u3068\u3067\uff0c\u8155\u90e8\u304c\u81ea\u7136\u306b\u5c55\u958b\u3059\u308b\uff08\u56f3 3-\u4e2d\uff09\uff0e \u8155\u90e8\u306f\uff0c\u6a5f\u4f53\u30d9\u30fc\u30b9\u90e8\u3068\u540c\u4e00\u5e73\u9762\u4e0a\u3068\u306a\u308b\u5730\u70b9\u3067\uff0c\u30b9\u30c8\u30c3\u30d1\u306b\u3088\u308a\u5c55\u958b\u304c\u5236\u9650\u3055\u308c\uff0c\u4e00\u822c\u7684\u306a\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u5f62\u72b6 \u3068\u306a\u308b\uff08\u56f3 3-\u53f3\uff09\uff0e\u8155\u90e8\u306e\u7a81\u8d77\u3068\u30b9\u30c8\u30c3\u30d1\u304c\u5d4c\u307e\u308a\u5408\u3046\u3053\u3068\u3067\uff0c\u30d2\u30f3\u30b8\u90e8\u306e\u904a\u3073\u306b\u3088\u308b\u8155\u90e8\u306e\u306d\u3058\u308c\u30fb\u632f\u52d5\u3092\u9632\u3050\uff0e\n\u306a\u304a\uff0c\u672c\u8ad6\u6587\u3067\u306f\u6298\u308a\u305f\u305f\u307f\u6a5f\u69cb\u306e\u6709\u7528\u6027\u3092\u691c\u8a3c\u3059\u308b\u3053\u3068\u3092\u76ee\u7684\u3068\u3059\u308b\u305f\u3081\uff0c\u6a5f\u4f53\u306b\u306f\u5b9f\u969b\u306e\u63a2\u67fb\u306b\u5229\u7528\u3059\u308b\n\u30bb\u30f3\u30b5\u3084\uff0c\u5916\u58c1\u3068\u306e\u63a5\u89e6\u9632\u6b62\u7528\u306e\u30d7\u30ed\u30da\u30e9\u30ac\u30fc\u30c9\u7b49\u306f\u642d\u8f09\u3057\u306a\u3044\uff0e\n2\u00b72\u00b71 \u8155\u90e8\u306e\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u306e\u691c\u8a3c\n\u8155\u90e8\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u3092\u691c\u8a3c\u3059\u308b\uff0e\u56f3 4\u306b\u53ef\u52d5\u8155\u90e8\u306b\u50cd\u304f\u529b\u3092\u793a\u3059\uff0e\u3053\u306e\u56f3\u306b\u304a\u3044\u3066\uff0c\u30d2\u30f3\u30b8\u306e\u56de\u8ee2\u4e2d\u5fc3\u3067\u3042\u308b\n\u70b9 Q\u56de\u308a\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u306e\u5f0f\u306f\u6b21\u5f0f\u306e\u3068\u304a\u308a\u3068\u306a\u308b\uff0e\u305f\u3060\u3057\uff0cl f \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u30d7\u30ed\u30da\u30e9\u4e2d\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0c l f g \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u53ef\u52d5\u8155\u90e8\u91cd\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0cm\u3092\u53ef\u52d5\u8155\u90e8\u306e\u8cea\u91cf\uff0cFm \u3092\u30d7\u30ed\u30da\u30e9\u63a8\u529b\uff0c\u03b8 f \u3092\u53ef\u52d5\u8155\u90e8\u306e \u6298\u308a\u305f\u305f\u307f\u89d2\u5ea6\uff0cg\u3092\u91cd\u529b\u52a0\u901f\u5ea6\u3068\u3059\u308b\uff0e\u306a\u304a\uff0c\u7c21\u5358\u306e\u305f\u3081\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306f\u8155\u90e8\u4e0a\u3067\u529b\u3092\u767a\u63ee\u3059\u308b\u3082\u306e\u3068\u3059\u308b\uff0e\n0 =\u2212l f gmgcos\u03b8 f + l f Fm (1)\n\u3053\u306e\u5f0f\u3092 Fm \u306b\u3064\u3044\u3066\u307e\u3068\u3081\u308b\u3068\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm = l f g\nl f mgcos\u03b8 f (2)",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3057\u3066\u3044\u308b\u5834\u5408\uff0c\u03b8 f = 0\u3067\u3042\u308b\u305f\u3081\uff0c\u53ef\u52d5\u8155\u90e8\u3092\u5b8c\u5168\u5c55\u958b\u72b6\u614b\u306b\u4fdd\u3064\u305f\u3081\u306e\u6761\u4ef6\u306f\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm \u2265 l f g\nl f mg (3)\n\u4e00\u65b9\u3067\uff0c\u6a5f\u4f53\u304c\u30db\u30d0\u30ea\u30f3\u30b0\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u6a5f\u4f53\u306e\u5168\u8cea\u91cf\u3092\u652f\u3048\u308b\u3060\u3051\u306e\u63a8\u529b\u304c\u5fc5\u8981\u3068\u306a\u308b\uff0e\u30db\u30d0\u30ea\u30f3\u30b0\u6642\uff0c4\u3064 \u306e\u30d7\u30ed\u30da\u30e9\u304c\u3059\u3079\u3066\u540c\u3058\u63a8\u529b\u3092\u767a\u63ee\u3057\u3066\u3044\u308b\u3068\u4eee\u5b9a\u3059\u308b\u3068\uff0c\u6a5f\u4f53\u5168\u4f53\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u304b\u3089\u30d7\u30ed\u30da\u30e9\u3042\u305f\u308a\u306e\u63a8\u529b Fm \u3092\u6c42\u3081\u308b\u3068\u6b21\u5f0f\u3067\u8868\u73fe\u3067\u304d\u308b\uff0e\u305f\u3060\u3057\uff0cmb \u3092\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u8155\u90e8\u4ee5\u5916\u306e\u6a5f\u4f53\u8cea\u91cf (\u30d9\u30fc\u30b9\u90e8\u8cea\u91cf)\u3068\u3059\u308b\uff0e\n4Fm = (mb +4m)g\nFm = ( 1 4 mb +m ) g (4)\n\u3053\u306e\u5f0f\u3068\uff0c\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3059\u308b\u305f\u3081\u306e\u6761\u4ef6\u5f0f (3)\u304b\u3089\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306b\u53ef\u52d5\u8155\u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u306a\u3044\u305f\u3081\u306e\u6761 \u4ef6\u304c\u6b21\u5f0f\u3068\u306a\u308b\uff0e( 1 4 mb +m ) g \u2265 l f g l f mg\nmb \u2265 4m ( l f g\nl f \u22121\n) (5)\n\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3067\u306f\uff0c\u53ef\u52d5\u8155\u90e8\u306e\u5148\u7aef\u306b\u6700\u91cd\u91cf\u7269\u306e\u30e2\u30fc\u30bf\u304c\u3064\u3044\u3066\u304a\u308a\uff0c\u305d\u306e\u70b9\u3067\u63a8\u529b\u304c\u767a\u63ee\u3055\u308c\u308b\uff0e\u305d\u306e \u305f\u3081\uff0cl f g < l f \u3068\u306a\u308a\uff0c\u5f0f (5)\u306e\u53f3\u8fba\u306f\u5e38\u306b\u8ca0\u3068\u306a\u308b\u3053\u3068\u304b\u3089\uff0c\u5e38\u306b\u6210\u308a\u7acb\u3064\uff0e\u3053\u308c\u3088\u308a\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u98db\u884c\u6642\u306b\u8155 \u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u3053\u3068\u306f\u7121\u3044\u3053\u3068\u304c\u8a3c\u660e\u3055\u308c\u305f\uff0e\n2\u00b72\u00b72 \u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u6642\u306e\u52d5\u4f5c\n\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306f\u3059\u3079\u3066\u306e\u30d7\u30ed\u30da\u30e9\u304c\u307b\u307c\u304a\u306a\u3058\u56de\u8ee2\u6570\u30fb\u63a8\u529b\u3092\u767a\u751f\u3057\u3066\u3044\u308b\u304c\uff0c\u79fb\u52d5\u306e\u305f\u3081 \u306b\u306f\u5404\u30d7\u30ed\u30da\u30e9\u3067\u63a8\u529b\u5dee\u3092\u4f5c\u308b\u3053\u3068\u3067\u6a5f\u4f53\u3092\u50be\u3051\uff0c\u6c34\u5e73\u65b9\u5411\u306e\u63a8\u529b\u3092\u5f97\u308b\uff0e\u307e\u305f\uff0c\u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u306f\uff0c\u5404\u30d7\u30ed\u30da\u30e9\u306b \u3088\u308a\u767a\u751f\u3059\u308b\u53cd\u30c8\u30eb\u30af\u3092\u5229\u7528\u3057\u3066\u5b9f\u73fe\u3059\u308b\uff0e\u305d\u306e\u305f\u3081\uff0c\u79fb\u52d5\u6642\u3084\u30e8\u30fc\u56de\u8ee2\u6642\u306b\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u3088\u308a\u3082\u63a8\u529b\u304c\u5c0f \u3055\u304f\u306a\u308b\u30d7\u30ed\u30da\u30e9\u304c\u5b58\u5728\u3059\u308b\u3053\u3068\u306b\u306a\u308a\uff0c\u524d\u8ff0\u306e\u8155\u90e8\u3092\u4fdd\u3064\u6761\u4ef6\u3092\u6e80\u305f\u3059\u3053\u3068\u304c\u3067\u304d\u306a\u3044\u5834\u5408\u304c\u5b58\u5728\u3059\u308b\uff0e\u7279\u306b",
            "\u6a5f\u4f53\u306e\u30e8\u30fc\u56de\u8ee2\u52d5\u4f5c\u306f\uff0c\u63a8\u529b\u306b\u5bfe\u3057\u3066\u5927\u5e45\u306b\u5c0f\u3055\u3044\u53cd\u30c8\u30eb\u30af\u3092\u7528\u3044\u308b\u305f\u3081\uff0c\u5404\u30d7\u30ed\u30da\u30e9\u3067\u306e\u56de\u8ee2\u6570\u5dee\u3092\u5927\u304d\u304f\u3059 \u308b\u5fc5\u8981\u304c\u3042\u308a\uff0c\u7279\u306b\u6ce8\u610f\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\n\u5177\u4f53\u7684\u306b\uff0c\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u52d5\u4f5c\u306f\uff0c\u56f3 5\u306e\u3088\u3046\u306b\u56de\u8ee2\u65b9\u5411\u306e\u7570\u306a\u308b\u30d7\u30ed\u30da\u30e9\u540c\u58eb\u306e\u56de\u8ee2\u6570\u306b\u5dee\u3092 \u3064\u3051\u308b\u3053\u3068\u3067\u5b9f\u73fe\u3059\u308b (Mahony et al., 2012)\uff0e\u4f8b\u3048\u3070\uff0c\u56f3 5(a)\u306b\u304a\u3044\u3066 CCW\u56de\u8ee2\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u56f3 5(b)\u306e\u3088\u3046\u306b CW\u56de\u8ee2\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u3092\u4e0a\u3052\uff0cCCW\u56de\u8ee2\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u3092\u4e0b\u3052\u308b\u3053\u3068\u3067\u5b9f\u73fe\u3055\u308c\u308b\uff0e\u3053\u306e\u6642\uff0c\u305d\u308c\u305e\u308c\u306e \u56de\u8ee2\u6570\u304c\u5909\u5316\u3057\u3066\u3082\uff0c\u5408\u8a08\u306e\u63a8\u529b\u304c\u5909\u5316\u3057\u306a\u3044\u3088\u3046\u306b\u5236\u5fa1\u3059\u308b\u3053\u3068\u3067\uff0c\u6a5f\u4f53\u306e\u9ad8\u5ea6\u3092\u4fdd\u3064\uff0e\u6a5f\u4f53\u306e\u5927\u304d\u306a\u30e8\u30fc\u89d2 \u52a0\u901f\u5ea6\u3092\u5f97\u308b\u305f\u3081\u306b\u306f\uff0c\u305d\u308c\u305e\u308c\u306e\u56de\u8ee2\u65b9\u5411\u306e\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u5dee\u3092\u5927\u304d\u304f\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\u307e\u305f\uff0c\u6a5f\u4f53\u306b\u50cd\u304f \u7a7a\u6c17\u62b5\u6297\u306b\u3088\u308a\uff0c\u9ad8\u901f\u30e8\u30fc\u56de\u8ee2\u3059\u308b\u5834\u5408\u306b\u306f\uff0c\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u5dee\u3092\u7dad\u6301\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\u3060\u304c\uff0c\u4e00\u822c\u306b\u30d7\u30ed\u30da\u30e9 \u306e\u56de\u8ee2\u6570\u5dee\u3068\u6a5f\u4f53\u306e\u30e8\u30fc\u56de\u8ee2\u901f\u5ea6\u306e\u95a2\u4fc2\u3092\u7406\u8ad6\u5024\u3088\u308a\u6c42\u3081\u308b\u3053\u3068\u306f\u56f0\u96e3\u3067\u3042\u308b\u305f\u3081\uff0c\u5b9f\u6a5f\u306b\u3088\u308a\u5b9f\u6e2c\u3057\u691c\u8a3c\u3059\u308b \u5fc5\u8981\u304c\u3042\u308b\uff0e\n2\u00b73 \u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u5b9f\u88c5\n\u63d0\u6848\u624b\u6cd5\u3092\u57fa\u306b\uff0c\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u5b9f\u88c5\u3057\u305f\uff0e\u56f3 6\u306b\u6a5f\u4f53\u306e\u5916\u89b3\u3092\uff0c\u8868 1\u306b\u6a5f\u4f53\u8af8\u5143\u3092\u793a\u3059\uff0e \u6a5f\u4f53\u306e\u6700\u5c0f\u5e45\u306f\uff0c\u5c55\u958b\u6642\u3067 618 mm\u3067\u3042\u308a\uff0c\u6298\u308a\u305f\u305f\u3080\u3068\u305d\u306e 48.5%\u7a0b\u5ea6\u3067\u3042\u308b 300 mm\u3068\u306a\u308a\uff0c\u5927\u5e45\u306b\u5c0f\u578b\u5316\n\u304c\u5b9f\u73fe\u3067\u304d\u305f\u3053\u3068\u304c\u308f\u304b\u308b\uff0e"
        ]
    },
    {
        "image_filename": "designv11_64_0002935_gt2016-56508-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002935_gt2016-56508-Figure4-1.png",
        "caption": "Figure 4: AFTB assembly",
        "texts": [
            " The positive modal damping indicates that the bearing is stable over a wide spectrum of external disturbance frequencies. As will be explained later along with experimental results, the modal analysis was further extended to conical mode to identify the source of bounded low frequency sub-synchronous vibrations observed in the experiment. The foils are made of 76.2\u03bcm thick Inconel 718; the AFB sleeve and AFTB back-plate are made of stainless steel. Bump foils and top foils are secured inside the AFB sleeve as shown in Figure 4. Each AFTB includes six sets of top foils and bump foils welded on the back-plate. The thrust bearing top foil is coated with 20\u03bcm thick Teflon \u00ae . The lift-off test is widely used to predetermine bearing static performance at low speeds. The test scheme is fundamentally similar to those described in [12, 13]; nevertheless, the hardware is much simpler since the use of this test is rather for bearing quality check routine. The test rig is driven by an electric motor that can reach 45krpm in 4 seconds"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003134_978-981-10-2875-5_70-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003134_978-981-10-2875-5_70-Figure7-1.png",
        "caption": "Fig. 7 Equivalent mechanism of the minimum composite unit",
        "texts": [
            " As similar as the above solving process, it can be known that the node C also has a translational DOF with respect to the node A. On the condition of considering the constraint influence of the second part of the tetrahedral element, the tetrahedral element can be equivalent to the mechanism shown in Fig. 6. It can be easily obtained that the number of the DOFs of the equivalent mechanism is one. According to the analysis in Sect. 2.2 the equivalent mechanism of the minimum composite unit shown in Fig. 2 can also be formulated, as shown in Fig. 7. Assuming that the foldable and deployable strut BG is removed from the equivalent mechanism shown in Fig. 7 and the joint connecting the node B to the node A is selected as the actuated joint, then the nodes C, D, E, F and G will move as the movement of the node B, as shown in Fig. 8. According to the characteristic of the DOF of the tetrahedral element there exist the following relations: AB1 AB Fig. 8 Configuration of the mechanism shown in this figure after movement From Eq. (6) it can be gotten that AB1 AB \u00bc AG1 AG , which will not be restricted after connecting the foldable and deployable strut BG to the nodes B and G"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002986_s00170-016-9507-2-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002986_s00170-016-9507-2-Figure1-1.png",
        "caption": "Fig. 1 Component models: a geometry model and b FE model using hexahedral elements",
        "texts": [
            " In general, the structural components of machine tools are constituted by cast iron using a casting molding method. It is well known that the material parameters of cast iron, such as elastic modulus and mass density, have not been determined. Thus, it is necessary to determine accurate material parameters when modeling a component made of cast iron. The mass density can be obtained by dividing the mass by the volume. However, the elastic modulus is not acquired as easily. In this study, a method based on the modal test is used to determine the elastic modulus of the component, which is shown in Fig. 1. The progress toward parameter determination is shown in Fig. 2. First, the experimental modal parameters of the component, such as natural frequencies and corresponding mode shapes, are obtained whenmodal testing is executed with freeto-free boundary conditions. Then, the FE model of this component is established using hexahedral elements, as shown in Fig. 1b. The elastic modulus values are in the range of 78.5 to 157 GPa, and Poisson\u2019s ratio ranges from 0.23 to 0.29. After a normal mode analysis, the simulated modal parameters can be obtained. To minimize the error associated with experimental and simulated natural frequencies, an accurate elastic modulus is obtained, which is shown in Table 1. As shown in Table 2, the largest error in the natural frequency under the same mode shape in both modal test results and the FE model analysis is less than 5 %, which illustrated the accuracy of material parameters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002481_s11071-016-2847-5-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002481_s11071-016-2847-5-Figure4-1.png",
        "caption": "Fig. 4 A free particle moving relative to an inertial frame F and a moving frame F\u0302",
        "texts": [
            " In the second example, motion of a constrainedparticle is investigated,with attention shifted to recognizingwhen amotion constraint is scleronomic or rheonomic. Finally, the classical problem of the rolling motion of a sphere over a rotating table is examined in the last example, by employing the set of scleronomicity conditions developed in Sect. 6. 9.1 Motion of a free particle relative to a moving frame Consider the spatial motion of a particle P of mass m subject to a known force f (t). The motion of this particle with respect to an inertial Cartesian coordinate frame F is described by three coordinates, x1, x2 and x3, as shown in Fig. 4. The configuration manifold is M = R \u00d7 R \u00d7 R \u2261 E3, with dimension n = 3, while the kinetic energy can be expressed in the form T = 1 2m[(v1)2 + (v2)2 + (v2)2], with vi = x\u0307 i (i = 1, 2, 3). Then, the corresponding metric matrix and the affinities are g = mI3 and \u039bk i j = 0, (136) respectively,where I3 is the 3\u00d73 identitymatrix.Moreover, the corresponding event manifold can be identified with the Cartesian product E\u03043 = E3 \u00d7 R, with metric matrix and affinities given by g\u0304 = [ mI3 0 0 g00 ] and \u039b\u0304K I J = 0 (I, J, K = 1, "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002225_powercon.2014.6993883-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002225_powercon.2014.6993883-Figure4-1.png",
        "caption": "Fig. 4. Dancing test machine main view",
        "texts": [],
        "surrounding_texts": [
            "POWERCON 2014 Paper No CP2836 Page 2/5\nB. Accurate identification and Tracking measurement (a) Get the original pattern sample of the measuring point First of all, pattern sample of the measured point should be obtained in the image. As shown in Fig. 2, the point to be measured is set and the image information red box as the measured points of the initial pattern sample is obtained.\n(b) Search and match pattern sample location in the image After the pattern of the measured point is obtained, it is\nc Matching calculation principle\nIn the image to be searched, a coordinate system such as (xs, ys) is established. Each pixel corresponds to a coordinate and also a strength of Is(xs, ys). Similarly, a pixel in the template also corresponds to a coordinate (xt, yt) and the intensity of It(xt, yt). Then, Diff(xs,ys,xt,yt) = | Is(xs,ys) \u2013 It(xt,yt) | is defined as the absolute difference of pixel intensity.\nrows cols\n0 0\n( , ) Diff( , , , ) T T\ni j SAD x y x i y j i j\n(1) In the images to be searched, the original template pattern\nmathematical expression can be explained as follows.\nrows cols\n0 0 ( , )\nS S x y SAD x y\n(2) Rows and columns of the image to be searched can be\nrepresented as Srows and Scols, and the Trows and Tcols represent the row and column of template pattern. In this way, the minimum SAD value of the matching degree is given an evaluation value. Finally, the position of the image is matched by this evaluating value.\nIII. Calibration method based on transmission line\ngalloping Through video tracking technology, point on the\ntransmission line is captured and tracked and the two-dimensional pixel variation curve of each point is got. Calibration is needed in order to get the actual distance corresponding to each pixel, so as to obtain the real motion trajectory of each feature point. The relationship between pixel value and actual distance is influenced by many factors such as camera site, lens angle, focal distance and so on. Therefore, in the process of every observation, position and the camera lens angle and focal distance should remain unchanged.\nA. Calibration method for transmission line feature based on monocular vision Calibration coefficient is calculated according to the basic\nprinciple of imaging. The main purpose of the calibration is that the video tracking results should be converted to the actual physical units.\nThe calibration coefficient K of vertical displacement at any point can be represented as:\ncosb DK f 3\nK-The calibration coefficient K of vertical displacement at any point\nb- The pixel size of CCD \u2014\u2014Camera angle, f\u2014\u2014The focal length of the lens, Unit: mm D\u2014\u2014Distance from the camera to the measuring surface,\nUnit: mm cos\n1000 fD d\n4 d\u2014\u2014Distance from the camera to the measuring surface,\nUnit: mm \u2014\u2014The measured relative to the angle between the optical axis of the camera 2 2 1\n0 0( ) ( ) tan bx x y y f 5\n1572 Session 2",
            "POWERCON 2014 Paper No CP2836 Page 3/5\nx,y\u2014\u2014The measured pixel coordinates, it is ensured When the measured point is selected, x0 y0\u2014\u2014Coordinate of center point the camera and can be queried, The calibration coefficient K of vertical displacement can be got through three formulas above. 2 2 1\n0 0cos ( ) ( ) tan 1000\ncos\nb fb d x x y y f\nK f 6\nIn the above formulas, only D and is unknown and require on-site measurement.\nB. Accuracy verification a Dancing test model validation\nIn order to verify the accuracy of the single camera, the contrast tests of measuring indoor is completed in the testing machine Zhejiang University civil engineering laboratory. Dancing test device has characteristics as follows: The device can simulate various frequencies and different amplitudes of conductor galloping through the wire vibration of different dynamic parameters. Fig 3 is the main view.\nTest object: Simulation of transmission lines. The measured point selection: as shown in the Fig. 5\nmeasured points were selected. Comparative method: the measured displacement of point 1 is known and the maximum amplitude is 200mm. The displacement of measured point 1 is measured using the method of analyzing the monocular vision and the results are shown in Fig. 6.\nMeasurement results based on video analysis method is 201.40mm. Compared to known swing 200mm, error is 0.7%.\nb Small test line verification\nSmall span test line was established by choosing suitable line parameters. Line structure parameters such as tension, sag and mitotic count can be adjusted in this small test line, and the shape and angle of the simulated ice can be replaced conveniently. Benchmarking traditional method can be used to get the galloping amplitude.\n1573 Session 2",
            "POWERCON 2014 Paper No CP2836 Page 4/5\nComprehensive test centre based on full-scale transmission line(\u2018CTC\u2019 for short), is the first comprehensive laboratory focus on ice-accretion and galloping of power transmission line, which is located on Jianshan Mountain, about 30km from southwest of Zhengzhou City in He\u2019nan province. CTC had been configured with full-scale test line, on-line monitoring system and monitoring station until the end of 2012. The dancing artificial simulation of double, four and six split transmission line with characteristic such as large amplitude, high frequency and long period is realized in the base.\nThis method is applied to the comprehensive test base of real transmission line to realize: (1) obtaining galloping amplitude, frequency and other parameters, and the effects of line structure, meteorological conditions and simulating ice installation angle. (2)getting the changes of galloping amplitude, frequency and other parameters with various anti device installed before and after and providing test method for the evaluation of the effect of anti-galloping device.\nThe comprehensive test base of real transmission line #3-#4 line dancing is taken as an example, the use of non-contact video analysis software to get the galloping amplitude, frequency of application, order is introduced.\nDancing line:#3-#4 test line The splitting number: Six Time: May 10, 2012 14:31 Conductor condition: install artificial icing Wind speed: 7.0m/s Status: anti-galloping measures is not installed Direction: South by Southwest Cameras: Canon 7D angle of altitude: 11.5 \u00b0 Angle with the line : 56 \u00b0\nThrough the above analysis it can be seen that the #3-#4 galloping vertical amplitude maximum value is 2.04 meters and the maximum amplitude value level is 0.63 meters. Wave frequency is 0.325Hz. The galloping order is two.\nTAB.2 EACH SPACER GALLOPING AMPLITUDE, FREQUENCY\nMeasure ment point The vertical displacemen t /m Horizontal displacemen t /m Horizontal frequency /Hz Vertical frequency\n/Hz 1 1.31 0.42 0.325 0.325 2 1.72 0.49 0.325 0.325 3 1.62 0.43 0.325 0.325 4 1.24 0.37 0.325 0.325 5 0.21 0.06 0.325 0.325 6 0.5 0.16 0.325 0.325 7 1.11 0.34 0.325 0.325 8 2.04 0.62 0.325 0.325 9 2.10 0.63 0.325 0.325\n10 1.91 0.59 0.325 0.325 11 1.32 0.4 0.325 0.325\n1574 Session 2"
        ]
    },
    {
        "image_filename": "designv11_64_0001169_1.a32416-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001169_1.a32416-Figure10-1.png",
        "caption": "Fig. 10 Cross-sectional model.",
        "texts": [
            " Results of the finite-element simulations indicated that the thickness of the cross section around the crease becomes larger than the other area in the cross section. The large thickness of the cross section induces higher compressive stress by wrapping the creased membrane, and thus the local buckling can be induced around the crease. Thus, the local buckling depends on the geometric properties of the cross section of the crease area. Based on the previous discussion, we introduce the cylindrical cross-section model, where the radius of the cylinder is equivalent to the layer thickness, as shown in Fig. 10. In the analysis of the local buckling, we have the following three steps. 1) Obtain the layer thickness to determine the radius of the cylindrical cross-section model a, where the deformation of the membrane in the creasing process is considered. 2) Examine the flattening of the cylindrical cross-section model due to the wrapping process, where we extend Aksel\u2019rad\u2019s buckling analysis [8,9] to derive the governing equations of the cylinder. 3) Identify the condition for the local buckling. In this step, the layer thickness of the creased membrane a is identified to determine the radius of the cylindrical model. As shown in Fig. 10a, the membrane is creased by the contact force from the center hub q. For thewrapped membrane, we assume that the contact area between the membrane and the center hub is sufficiently small, and hence we treat the contact force as the concentrated force. In our previous research, a layer thickness of creasedmembrane induced by a concentrated force is obtained [10] as a \u2243 Et3 24q s c1; c1 Z \u03c0\u22152 0 cos \u03c6 p d\u03c6 \u2243 1.2 (1) whereE, t, and\u03c6 are Young\u2019s modulus, themembrane thickness, and the variable of integration, respectively",
            " The circular dots and the triangular dots represent the experimental and the FEMresults,where the solid line and the broken line show results of the creaseCL andCR, respectively. Additionally, the square dots represent the results of the theoretical analysis. As shown in Fig. 17, the interval of local buckling ismonotonously increased with the radius of center hub, and the results of the theoretical analysis show qualitatively good agreement with experimental and FEM results. When the results are evaluated quantitatively, a larger interval is obtained using theoretical analysis than in the experimentally obtained results. As shown in Fig. 10, larger curvature \u03ba1 is evaluated when we introduce the cylindrical cross-section model. As shown in Eq. (23), the larger curvature \u03ba1 induces a larger interval of local buckling lb. Therefore, the difference of the interval between the results of the theoretical analysis and the experiments are caused mainly by the difference of the curvature \u03ba1 when the cylindrical model is introduced. The mechanics of local buckling induced in a crease, which determines the folded size and the residual deformation of space membranes, was identified in consideration of the deformed shape of the creased membrane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001943_wzee.2015.7394035-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001943_wzee.2015.7394035-Figure1-1.png",
        "caption": "Fig. 1. Cross section of a motor with marked measurement points, where: 1\u03d1 \u2013 temperature between the stator and motor housing, 2\u03d1 \u2013 temperature of the motor housing between fins, 3\u03d1 \u2013 fin temperature, o\u03d1 \u2013 ambient temperature",
        "texts": [
            " The proposed method of determining this value allows for correct determining of the \u03b1 factor for the particular motor housing at some characteristic point of the body. Additionally, it allows to determine the dependency of the \u03b1 factor on the motor revolution speed, but not the cooling medium velocity, which undoubtedly is more practical during implementation in mathematic models and more intuitive for each designer. According to Newton\u2019s condition the heat flux q\u03b1 conducted in solid matter is equal to the flux transferred to the environment. ( ) oF Fn q \u03d1\u03d1\u03b1\u03d1\u03bb\u03b1 \u2212= \u2202 \u2202= (1) Performing temperature measurement in the points indicated in Figure 1 it is possible to determine the factor \u03b1. On the basis of Fourier\u2019s law and knowledge of temperatures 321 \u03d1\u03d1\u03d1 , , as well as precise dimensions of the housing, it is possible to determine the heat flux conducted by the housing. Knowing the surface temperature of a structure and of the environment, on the basis of the relation (1) it is possible to determine the heat transfer coefficient between the fins (2) and for the fins themselves (3). ( ) 2 2 2 1 1 21 o o ll \u03d1\u03d1 \u03bb\u03bb \u03d1\u03d1\u03b1 \u2212\u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b + \u2212= (2) where: l1, (l2) \u2013 thickness of air layer (housing), \u03bb1, (\u03bb2) \u2013 air (aluminium) heat conductivity factor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure16-1.png",
        "caption": "Fig. 16 Force components affecting internal normal force in a femur, b tibia\u2014muscles considered",
        "texts": [
            " As muscle insertion points do not lay on the bone\u2019s neutral axis, x components of the muscle force will also create point moments at the insertion points. Due to the parametric nature of the model, the order at which muscle and gravity forces are applied along the bone can vary. Therefore, the algorithm for the computation of those forces will be presented instead of a particular case solution. At maximum, four sections can be identified. However, due to the assumption of no co-contraction, one of the muscle forces will always be zero, reducing the effective number of zones to three. Normal force components are presented in Fig. 16. The new force component at the right side of the femur can be seen with w index; this represents the reaction force from the muscle wrapping points. Similarly to model A1, normal forces in the bone are additive. To find the value of the normal force in a segment, all forces depicted in Fig. 16 preceding this segment should be summed up. The + = operator is used to indicate the summation of the components that fulfill the conditional requirements. N1+ = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 Fhx1 x1 \u2265 0 Fmex1 x1 \u2265 I\u03011x L1 m1gx1 x1 \u2265 Cm1L1 Fm f x1 x1 \u2265 I\u03011bx L1 (47) N2+ = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 Fkx2 x2 \u2265 0 Fmex2 x2 \u2265 I\u03012x L2 m2gx2 x2 \u2265 Cm2L2 Fm f x2 x2 \u2265 I\u03012bx L2 (48) Shear forces are calculated according to the same principle as normal forces, by adding force components that fulfill the x coordinate condition. The shear force components are depicted in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.8-1.png",
        "caption": "FIGURE 6.8",
        "texts": [
            " This model is developed from the lumped mass model by treating the front and rear suspensions as rigid axles connected to the body by revolute joints. The locations of Equivalent roll stiffness model approach. REV, revolute joint. the joints for the two axles are their respective \u2018roll centres\u2019 as described in Chapter 4. A torsional spring is located at the front and rear roll centres to represent the roll stiffness of the vehicle. The determination of the roll stiffness of the front and rear suspensions required an investigation as described in the following section. The equivalent roll stiffness model is shown schematically in Figure 6.8. Note that this model shows the historical background to much of the current unclear thinking about roll centres and their influence on vehicle behaviour. With beam axles, as were prevalent in the 1920s, this model is a good equivalent for looking at handling behaviour on flat surfaces and ignoring ride inputs. For independent suspensions where the anti-roll geometry remains relatively consistent with respect to the vehicle and where the roll centres are relatively low (i.e. less than around 100 mm for a typical passenger car) e a fairly typical double wishbone setup, for example e then this approximation can be useful despite its systematic inaccuracy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001615_1.4031579-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001615_1.4031579-Figure3-1.png",
        "caption": "Fig. 3 Definition of seal compression",
        "texts": [
            " For this model, it is assumed that seal stiffness is affected by three contributions: the stiffness force fS, which arises due to the mechanical forces that need to be overcome in order to compress the filaments, and two aerodynamic contributions, namely, the pressure force fP and inertia force fI (described in more detail as part of Sec. 3). For small changes in seal pack geometry, as may arise during rotor excursions and rotor diameter changes (less than 1 mm), it is possible to approximate the forces acting on a single filament by the following linearized relationships: fS \u00bc p1 Dx\u00fe p0 (1) fP \u00bc q2 Dx DP\u00fe q1 DP\u00fe q0 \u00bc DP\u00f0q2 Dx\u00fe q1\u00de (2) fI \u00bc r2 Dx DP\u00fe r1 DP\u00fe r0 \u00bc DP\u00f0r2 Dx\u00fe r1\u00de (3) with units Newtons (per leaf). Here Dx is the distance by which the seal filaments have been compressed as shown in Fig. 3 and p, q, and r are seal-specific constants used in the linearized models. For the stiffness force, Eq. (1), p1 is representative of the filament stiffness and p0 is an offset that may arise due to uncertainty in the nominal seal bore position. For the pressure and inertia force, Eqs. (2) and (3), respectively, q0 and r0 are constants that can be set to zero as fP and fI are zero for the no-flow case (DP \u00bc 0). q1 and r1 correspond to constants that capture the relationship between tip force and flow through the seal (by considering DP)",
            " The total force exerted on the rotor by a single filament is obtained by summing the three contributions fT \u00bc fS \u00fe fP \u00fe fI (4) 011004-2 / Vol. 138, JANUARY 2016 Transactions of the ASME Downloaded From: http://turbomachinery.asmedigitalcollection.asme.org/ on 01/27/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use This total force can be used to calculate either seal torque or seal stiffness depending on how the force at the rotor interface is resolved. The seal compression Dx is a function of local rotor diameter, seal bore, and also seal blow-down as shown in Fig. 3. Here RR is the position of the rotor surface relative to the origin, given by RR \u00bc R0 \u00fe Dr|{z} radial growth \u00fe d cos h|fflfflffl{zfflfflffl} rotor offset (5) Thus for seal arrangements in which contact between the elements and the rotor is maintained at all times, either through sufficient cold build interference or blow-down, the expression for Dx simplifies to Dx \u00bc Dr \u00fe d cos h (6) where h is the nominal angular position given by h \u00bc n\u00f02p=N\u00de and n \u00bc 1; 2; 3; :::N. This expression will be used in further analyses"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001391_dscc2014-6301-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001391_dscc2014-6301-Figure2-1.png",
        "caption": "Figure 2. Conventional steering system",
        "texts": [
            " \u2212 C3 \u03b1 27\u00b52F2 z tan3 \u03b1, |\u03b1|< tan\u22121 ( 3\u00b5Fz C\u03b1 ) \u2212\u00b5Fzsgn \u03b1, otherwise (2) = ftire (\u03b1) (3) where \u00b5 is the surface coefficient of friction, Fz is the normal load, and C\u03b1 is the tire cornering stiffness. The tire slip angles, \u03b1f and \u03b1r, can be derived from the kinematics of the vehicle as: \u03b1f = tan\u22121 ( \u03b2+ ar Ux ) \u2212\u03b4 (4) \u03b1r = tan\u22121 ( \u03b2\u2212 br Ux ) (5) where \u03b4 is the steer angle. The vehicle\u2019s position is specified relative to a reference line using three states: heading deviation \u2206\u03c8, lateral deviation e, and distance along the path s as defined in Figure 1. The equations of motion of these states can be written as: \u2206\u0307\u03c8 = r (6) e\u0307 = Ux sin(\u2206\u03c8)+Uy cos(\u2206\u03c8) (7) s\u0307 = Ux cos(\u2206\u03c8)\u2212Uy sin(\u2206\u03c8) (8) Figure 2 illustrates the mechanical coupling of the handwheel and the roadwheels in a conventional steering system. In steer-by-wire vehicles, the handwheel is mechanically decoupled from the roadwheels. A force feedback (FFB) steering system can be used in tandem with a steering feedback model to create realistic steering feedback for the driver on a steer-by-wire vehicle. Balachandran and Gerdes [7] proposed a steering feedback model with sufficient complexity to capture the important elements of steering feedback that drivers care about while still being implementable in real-time on a steer-by-wire vehicle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.11-1.png",
        "caption": "Fig. 3.11 Formulation of the constitutive law based on a stress and b stress resultant",
        "texts": [
            "20) is not so easy to apply12 in the case of beams since the stress and strain is linearly changing over the height of the cross section, see Eq. (3.26) and Fig. 3.9. Thus, it might be easier to apply a so-called stress resultant or generalized stress, i.e. a simplified representation of the normal stress state13 based on the acting bending moment: My(x) = \u222b\u222b z\u03c3x(x, z) dA, (3.28) which was already introduced in Eq. (3.22). Using in addition the curvature14 \u03ba = \u03ba(x) (see Eq. (3.16)) instead of the strain \u03b5x = \u03b5x(x, z), the constitutive equation can be easier expressed as shown in Fig. 3.11. The variables My and \u03ba have both the advantage that they are constant for any location x of the beam. 12However, this formulation works well in the case of rod elements since stress and strain are constant over the cross section, i.e. \u03c3x = \u03c3x(x) and \u03b5x = \u03b5x(x), see Fig. 2.4. 13A similar stress resultant can be stated for the shear stress based on the shear force: Qz(x) =\u222b\u222b \u03c4xz(x, z) dA. 14The curvature is then called a generalized strain. 3.2 Derivation of the Governing Differential Equation 101 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001968_s40435-016-0225-2-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001968_s40435-016-0225-2-Figure1-1.png",
        "caption": "Fig. 1 Schematic of the biped robot model on a ramp",
        "texts": [
            " First, a normalized target energy is chosen and then, the control law is determined to incline the normalized energy of the robot toward the normalized target energy. Then, using the biped model proposed byWisse et al. [8], the biped robot is controlled by applying the proper angular movement to the robot\u2019s upper body. It is shown that the robot can be controlled by exerting proper movements to its upper body and its instance leg and the proposed active control schemes can improve the robot\u2019s stability. In this research a simple 2D model of a kneeless biped robot proposed by Wisse et al. [8] is used as shown in Fig. 1. In this model all masses are considered as points. The counter clockwise direction is considered as positive direction for angles. Besides, it is assumed that walking has two phases: swing phase, swing around a simple support and the transition phase, instantaneous change of the stance leg. At swing phase, the robot acts like as a double planar pendulum. At transition phase, the support transmits from one leg to the other one immediately. In this paper, it is assumed that the robot moves in a plane and is restricted from falling from sides",
            " Let denote following parameters for brevity: \u03b2 = m M , \u03bc = M1 M , \u03ba = l1 l , \u03be = \u03b2 1 + \u03bc , \u03b7 = \u03bc\u03ba 1 + \u03bc , (1) where M is pelvis mass, M1 is upper body mass, l is leg length, l1 is body length, and m is foot mass and scale time by \u221a g l . Then the swing equations are simply obtained by energy method as: [ 1 + \u03be (1 \u2212 cos (\u03c6)) \u2212\u03be cos (\u03c6) 1 \u2212 cos (\u03c6) 1 ] [ \u03b8\u0308 \u03c6\u0308 ] + sin (\u03c6) [ \u03be (\u03c6\u0307 + \u03b8\u0307 )2 \u2212\u03b8\u03072 ] + [\u2212(1 + \u03be) sin (\u03b8 + \u03b3 ) sin (\u03c6 + \u03b8 + \u03b3 ) ] = \u2212\u03b7 [ cos (\u03b8 \u2212 \u03b1) \u03b1\u0308 + sin (\u03b8 \u2212 \u03b1) \u03b1\u03072 0 ] (2) where \u03b8 is the absolute stance leg angle, \u03c6 is the relative swing leg angle, \u03b1 is the absolute upper body angle, and \u03b3 is the slope angle shown in Fig. 1. The geometrical constraint of collision is \u03c6 = \u22122\u03b8 . Right after of the collision, one has the below relations: \u03b8+ = \u2212\u03b8\u2212 := \u03b8 \u03d5+ = \u2212\u03d5\u2212 = \u22122\u03b8 \u03b1+ = \u03b1\u2212 := \u03b1 (3) Since there is no strike reaction when the rear leg leaves the ground, it can be assumed the angular momentum of the rear leg, the prior stance leg and the future swing leg, is constant around the waist. Beside this, the angular momentum of upper body around the waist remains constant during the collision since there is no strike exerted to the upper body",
            " Despite the increase of \u03bb causes an increase in the range of slopes having stable gaits, but on the other hand the increase of \u03bb means an increase in the exerted energy to the system without changing the walking speed, therefore it decreases the efficiency of gaits. 11. For small \u03bbs, increase of the normalized target energy has less effect on the system. Whereas for larger \u03bbs, the same increase in the normalized target energy causes more changes in the system. In this section, control of a bipedal walking robot shown in Fig. 1 is considered by exerting proper active movements on its upper body. Study of this model is helpful to understand the effect of bipedal walking robot\u2019s upper body on walking and gait stability. The control method used in this section is based on energy pass tracking. In fact it is tried to reach the energy level of legs to the desired energy level by applying the propermovements to the upper body. It can be seen that all terms related to the upper body angle, \u03b1, are at the right side of Eq. (2)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.32-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.32-1.png",
        "caption": "Fig. 3.32 Beam element with rotated support: a rotation at node 1; b rotation at node 2",
        "texts": [
            " E \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 12 I L 3 si n2 \u03b1 + A L co s2 \u03b1 ( 12 I L 3 \u2212 A L ) si n\u03b1 co s\u03b1 \u2212 6I L 2 si n\u03b1 \u2212 12 I L 3 si n2 \u03b1 \u2212 A L co s2 \u03b1 ( \u2212 12 I L 3 + A L ) si n\u03b1 co s\u03b1 \u2212 6I L 2 si n\u03b1 ( 12 I L 3 \u2212 A L ) si n\u03b1 co s\u03b1 12 I L 3 co s2 \u03b1 + A L si n2 \u03b1 \u2212 6I L 2 co s\u03b1 ( \u2212 12 I L 3 + A L ) si n\u03b1 co s\u03b1 \u2212 12 I L 3 co s2 \u03b1 \u2212 A L si n2 \u03b1 \u2212 6I L 2 co s\u03b1 \u2212 6I L 2 si n\u03b1 \u2212 6I L 2 co s\u03b1 4I L 6I L 2 si n\u03b1 6I L 2 co s\u03b1 2I L \u2212 12 I L 3 si n2 \u03b1 \u2212 A L co s2 \u03b1 ( \u2212 12 I L 3 + A L ) si n\u03b1 co s\u03b1 6I L 2 si n\u03b1 12 I L 3 si n2 \u03b1 + A L co s2 \u03b1 ( 12 I L 3 \u2212 A L ) si n\u03b1 co s\u03b1 6I L 2 si n\u03b1 ( \u2212 12 I L 3 + A L ) si n\u03b1 co s\u03b1 \u2212 12 I L 3 co s2 \u03b1 \u2212 A L si n2 \u03b1 6I L 2 co s\u03b1 ( 12 I L 3 \u2212 A L ) si n\u03b1 co s\u03b1 12 I L 3 co s2 \u03b1 + A L si n2 \u03b1 6I L 2 co s\u03b1 \u2212 6I L 2 si n\u03b1 \u2212 6I L 2 co s\u03b1 2I L 6I L 2 si n\u03b1 6I L 2 co s\u03b1 4I L \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 u 1 X u 1 Z \u03d5 1Y u 2 X u 2 Z \u03d5 2Y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 F 1X F 1Z M 1Y F 2X F 2Z M 2Y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 (3 .2 94 ) 3.4 Assembly of Elements to Plane Frame Structures 159 160 3 Euler\u2013Bernoulli Beams and Frames It might be required for certain problems to apply the transformation given in Eq. (3.293) only at one node of the element. This can be the case if a support is rotated only at one node as shown in Fig. 3.32. Let us have a look, for example, on the case shown in Fig. 3.32awhere it would be quite difficult to describe the boundary condition in the local (x, z) system. However, the global (X,Z) system easily allows to specify the boundary conditions at the first node as: u1Z = 0 \u2227 u1X = 0. Thus, the transformation (3.293) can be individually applied at each node with a different transformation angle \u03b11 at node 1 and \u03b12 at node 2 as: T = [ T1(\u03b11) 0 0 T2(\u03b12) ] . (3.295) The last equation implies that the global coordinate system can be differently chosen at each node. In the case that the rotation is only required at the first node as shown in Fig. 3.32a, Eq. (3.295) can be simplified to T = [ T1(\u03b11) 0 0 I ] , (3.296) where I is the identity matrix. Thus, the elemental stiffness matrix in the global coordinate system can be obtained for this special case based on the following relationship: 3.4 Assembly of Elements to Plane Frame Structures 161 Ke XZ = [ T1(\u03b11) 0 0 I ]T Kxz [ T1(\u03b11) 0 0 I ] . (3.297) In a similar way, the transformation can be only performed at node 2. The elemental stiffness matrices for these two special cases are summarized in Eqs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001494_amm.732.357-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001494_amm.732.357-Figure3-1.png",
        "caption": "Fig. 3 Model of plate loading",
        "texts": [
            " From these determined quantities of the global system (x, y z), which will be transformed to the local system ( )321 ,, n , we determine ( ) \u03b5T\u03b5 nn L n \u03b8= , ( ) \u03b3T\u03b3 t nn L n \u03b8= . (20) The transformation matrixes [ ]( )\u03b8ntTT, are specified by Eq. 4. Eq. 1 is applied for evaluation stress in n-th layer. We note that, the calculation based on the Mindlin-Reissner theory gives incorrect results of the stress in the vicinity of the free edges. These values can be used as the initial input for the stress determination by more accurate procedure. The square plate 250 \u00d7 250 \u00d7 0.89 mm (Fig. 3) was solved by described method. The plate is unidirectional composite consisting of carbon fibers embedded in the epoxy matrix. The fibers are deposited in the matrix parallel to one edge of the plate. The direction of the fibers is denoted as direction \u201c11\u201d. Other directions are denoted \u201c22\u201d and \u201c33\u201d. The density of the plate is 1 540 kg\u22c5m-3. Applied Mechanics and Materials Vol. 732 361 The material constants are: Young\u2019s modulus in the particular direction E11 = 1.299\u22c51011 Pa, E22 = 1.39\u22c51010 Pa, E33 = E22, Poisson\u2019s ratio \u00b512 = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001481_1077546314538480-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001481_1077546314538480-Figure2-1.png",
        "caption": "Figure 2. Each finite element of two members in local system of coordinates.",
        "texts": [
            " The roots of this equation are the frequency parameters of the system, i.e., l. In solving the frequency equation, the following relations should be employed: l \u00bc ffiffiffiffiffiffiffiffiffiffiffiffi EI E I A A 4 s l \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffi I0EI AGJ s l2 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffi I0 E I A G J s l2 8>>>>>>>>>>>< >>>>>>>>>>>: \u00f016\u00de In this section, a finite element formulation of the problem is presented. The nodal displacements with respect to the local systems of coordinates are identified in Figure 2. It is interesting to note that the torsional stiffness is combined with the flexural stiffness matrix to obtain the stiffness matrix for a typical member of a grid structure. Furthermore, the combination of the consistent mass matrix for flexural effects with the torsional mass matrix for torsional effects results in the consistent mass matrix for a typical member of a grid (Paz, 1985). The stiffness and consistent mass matrices of the first member are \u00bdKe \u00bc EI L3 12 6L 0 12 6L 0 6L 4L2 0 6L 2L2 0 0 0 GJL2=EI 0 0 GJL2=EI 12 6L 0 12 6L 0 6L 2L2 0 6L 4L2 0 0 0 GJL2=EI 0 0 GJL2=EI 0 BBBBBBB@ 1 CCCCCCCA \u00f017\u00de \u00bdMe \u00bc AL 420 156 22L 0 54 13L 0 22L 4 0 13L 3L2 0 0 0 140I0= A 0 0 70I0= A 54 13L 0 156 22L 0 13L 3L2 0 22L 4L2 0 0 0 70I0= A 0 0 140I0= A 0 BBBBBBB@ 1 CCCCCCCA \u00f018\u00de at Ondokuz Mayis Universitesi on November 6, 2014jvc"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002283_edpc.2014.6984426-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002283_edpc.2014.6984426-Figure4-1.png",
        "caption": "Fig. 4. Dynamic model of a helical gear pair system",
        "texts": [
            " The frequency response shows a slight positive gain in the passband region of about +1 dB. For ANC applications the system input Tdes(t) = Tstat(t) + \u0394T(t) (2) consists of a static or low bandwidth component Tstat(t) representing the torque input e.g. from a speed controller as depicted in Fig. 2 and a high bandwidth alternating signal component \u0394T(t) representing the ANC output. The torsional mesh dynamics of the helical gear pair has been developed by Tuplin [2] and used by many other researchers [3], [6], [7]. Fig. 4 shows the torsional dynamic model of a helical gear pair considered in this paper. Subscript P indicates the pinion parameters and subscript G the gear parameters. The model consists of two rigid bodies each with a lumped moment of inertia. The coupling of both gears is modeled with a spring-damper element where the spring stiffness ( ) ( ))cos(10 PPPm kk \u03d5\u03bd\u03b5\u03d5 += , (3) with the gear mesh coefficient \u03b5, is dependent on the pinion angle \u03d5P, which provides a nonlinearity in the equations of motion"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003101_icelmach.2016.7732702-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003101_icelmach.2016.7732702-Figure1-1.png",
        "caption": "Fig. 1. Studied machine: (a) Fractional-slot STPM machine (6-slots/4-poles) with the buried PMs and the single layer winding (i.e., non-overlapping alternate teeth wound winding), and (b) Definition of regions for the analytial model.",
        "texts": [
            " This model with single layer winding is valid for any number of slot/pole combinations and can be extended easily for double layer winding. The 2-D analytical results considering the drive current (i.e., sinusoidal and six-step rectangular current) are compared with those obtained by 2-D Fem [16]. The comparisons (viz., the electromagnetic performances and the non-intrinsic UMF) are very satisfying in amplitudes and waveforms. II. STUDIED MACHINE AND MAGNETIC FIELD SOLUTION A. Problem Description and Assumptions Fig. 1 shows the fractional-slot STPM machine where Region I represents the air-gap, Region II the buried PMs, Regions III the stator slots, Region IV the non-magnetic material under PMs. The 2-D analytical model is formulated in magnetic vector potential and polar coordinates with the following assumptions: - The axial length of the machine is infinite and invariant (i.e., end-effects are neglected); - The iron parts are assumed to be infinitely permeable; - The rotor and stator tooth-tips are not considered"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001095_icra.2014.6907630-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001095_icra.2014.6907630-Figure3-1.png",
        "caption": "Fig. 3. Illustration of one learning iteration. On the timeline (top), the vehicle is shown performing the open-loop maneuver, followed by the recovery back to hover. Two estimates of the final state s(tf ) are computed for each iteration: one by forward integration from time t0 to tf , and one by backward integration from time tr to tf . Subsequently, the two estimates are fused and used to perform a parameter adaptation, reducing the expected value of the next final state deviation.",
        "texts": [
            " Doing so, the step size \u03b3i does not converge to zero, meaning that the learning algorithm is still adaptive after a large number of iterations. After each parameter update, we also have to recompute the trajectories of the desired total thrust force fdestot and of the desired body rates \u03c9des, as these are the inputs fed to the onboard controller when the open-loop maneuver is executed. We now address the problem of estimating the final state offset \u2206sf by observing one execution of the open-loop maneuver. Note that for reasons of readability the superscript i, denoting to current learning iteration, is omitted in the following. Fig. 3 shows the different intervals of such a learning iteration: The vehicle executes the high-performance open-loop maneuver between time t0 and tf , and recovers to hover between time tf and tr. During recovery, the vehicle receives control commands from an external pilot that can take advantage of absolute position measurements. We therefore assume that the vehicle is guided back to the position from where it started the open-loop maneuver. 1) Forward State Integration: We assume that the vehicle is hovering before the maneuver is triggered; the expected initial state s\u03020 is known, and the initial covariance matrix \u03a3[s0] is a design parameter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000589_pc.2015.7169967-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000589_pc.2015.7169967-Figure1-1.png",
        "caption": "Fig. 1. Schematic of the pendubot in a relative coordinate system.",
        "texts": [
            " The generalized coordinates summarized in the vector q = [q1, q2]T here stand for the angular positions of the two links, and \u03c4 = [\u03c41,0]T denotes the external control force vector. By applying (1), the resulting equation of motion can be cast in the standard vector/matrix form: D(q)q\u0308 +C(q, q\u0307)q\u0307 +F (q\u0307) + g(q) = \u03c4 (2) where D(q) is the symmetric positive definite inertia matrix, C(q, q\u0307) contains the Coriolis and centrifugal terms, F (q\u0307) is the vector of viscous frictional terms, and g(q) denotes the vector of gravitational terms. For the Pendubot system, schematically illustrated in Fig. 1, the following quantities can be obtained: D(q) = [\u03b81 + \u03b82 + 2\u03b83 cos q2 \u03b82 + \u03b83 cos q2 \u03b82 + \u03b83 cos q2 \u03b82 ], C(q, q\u0307) = [\u2212\u03b83 sin(q2)q\u03072 \u2212\u03b83 sin(q2)q\u03072 \u2212 \u03b83 sin(q2)q\u03071 \u03b83 sin(q2)q\u03071 0 ], g(q) = [\u03b84g cos q1 + \u03b85g cos(q1 + q2) \u03b85g cos(q1 + q2) ] , where \u03b81 =m1l 2 c1 +m2l 2 1 + I1 \u03b82 =m2l 2 c2 + I2 \u03b83 =m2l1lc2 \u03b84 =m1lc1 +m2l1 \u03b85 =m2lc2 are the parameter equations needed for the subsequent control design. Taking into account viscous friction in both joints the dynamic system (2) can be rewritten as follows: q\u03081= 1 \u03b81\u03b82\u2212\u03b823cos 2q2 [\u03b82\u03b83sinq2(q\u03071+q\u03072)2+\u03b823cosq2sin(q2)q\u030721 \u2212\u03b82\u03b84gcosq1+\u03b83\u03b85gcosq2cos(q1+q2)+\u03b82\u03c41\u2212\u03b82b1q\u03071 +(\u03b82+\u03b83cosq2)b2q\u03072] (3) q\u03082= 1 \u03b81\u03b82\u2212\u03b823cos 2q2 [\u2212\u03b83(\u03b82+\u03b83cosq2)sinq2(q\u03071+q\u03072)2 \u2212(\u03b81+\u03b83cosq2)\u03b83sin(q2)q\u030721+(\u03b82+\u03b83cosq2)(\u03b84gcosq1\u2212\u03c41) \u2212(\u03b81+\u03b83cosq2)\u03b85gcos(q1+q2)+(\u03b82+\u03b83cosq2)b1q\u03071 \u2212(\u03b81+\u03b82+2\u03b83cosq2)b2q\u03072]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001258_j.wear.2015.01.039-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001258_j.wear.2015.01.039-Figure3-1.png",
        "caption": "Fig. 3. Photographs of: (a) the reciprocating linear pin-on-plane tribometer called the macrotribometer and (b) the used slider.",
        "texts": [
            " The forces tangential and normal to the sliding direction, respectively Ft and Fn, are measured in order to determine the total friction force F, the vector sum of the normal and tangential forces, and m, the friction coefficient (COF) (Fig. 2). Two different experimental devices at two different force scales have been used in order to study the behaviour of a single bristle or loop as well as more complex loop assemblies. A macrotribometer is used for the single loop and loop assemblies (surface with macropile) and a nanotribometer is used for the single bristle. In fact, the forces are not in the same range for a loop assembly and a single bristle. The macrotribometer is a reciprocating linear pin-on-plane tribometer (Fig. 3a) [22]. It is composed of an oscillating motorcontrolled table (Controls Linear Table x.act LT 100-2 ST controlled by a three-channel Stepper-Motor Controller M50.PCI, LINOS Photonics GmbH & Co. KG, G\u00f6ttingen, Germany), onto which the sample is affixed. A computer programme written in Visual Basic allows the oscillating table to have a reciprocating linear movement. The slider Table 1 Nomenclature of the loops samples. Designation Schema Number of columns of loops Number of rows of loops Interval between loops (mm) Column direction Row direction 1 1 1 1 \u2013 \u2013 1 7 1 7 \u2013 3 2 7 2 7 10 3 3 7 3 7 3 3 is mounted on an arm fixed to the frame with a pivot",
            " The normal load depends on the normal action of the sample on the slider and is measured with a piezoelectric force sensor with a charge amplifier (8200 and 2628, Br\u00fcel & Kjaer, Mennecy, France). In addition to the force sensors, a laser sensor (Micro-Epsilon OptoNCDT 670 nm) measures the table displacement during friction experiments. An image acquisition system (AVT PIKE F-032B) is also placed tangentially in order to observe the contact between the slider and the pile. The slider is cuboid shaped (Fig. 3b). It is made of aluminium with a roughness Ra\u00bc0.1870.01 mm. Its width is around 22 mm. The nominal sliding velocity ranged from 1 to 20 mm/s in order to correspond to the range of velocities when a person touches a textile surface. The second experimental device, only used for investigating the behaviour of a single bristle, is a NTR2 nanotribometer (CSM Instrument Company, Peseux, Switzerland). This device is a pinon-disk tribometer with reciprocating movement allowed (Fig. 4). A slider is fixed to a cantilever, which allows the measurement of the normal and tangential forces, Fn and Ft, respectively, thanks to capacitive sensors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure1-1.png",
        "caption": "Figure 1 Deposition configuration of fused deposition modelling (see online version for colours)",
        "texts": [
            " It has attracted extensive attention in recent years. The technology involves creating objects through sequential layering. Fused deposition modelling (FDM) is one of the most popular AM technologies. In this process, partially melted filament is extruded from a nozzle and deposited onto a platen to form objects (Vaezi, 2013). The motion of the nozzle is controlled by computer software from 3D model data to deposit material strand by strand and layer upon layer. The schematic of FDM is shown in Figure 1. Comprehensive investigations have been carried out in order to optimise the FDM process and improve the quality and accuracy of final products. Ann et al. (2002) and Sood et al. (2010, 2012) examined the influences of process parameters, such as raster orientation, air gap, bead width, layer thickness, and temperature on mechanical properties of final products. They also compared the tensile/compressive strength of FDM specimen with that of specimen made with injection moulding. Yardimci et al",
            " Predictions from the models were compared with experimental results, and a good agreement was achieved between the predicted and experimental data. The presented work can aid engineers in designing and analysing parts made from FDM. In order to characterise FDM material, torsion test and tensile test were conducted to obtain the elastic constants in the constitutive model. All of the specimens were made of ABS material with Young\u2019s modulus 2480 MPa. All the tests followed ASTM standard. FDM specimen can be considered as homogeneous and linear elastic orthotropic material. The 1, 2 and 3 axes as shown in Figure 1 are the three mutually orthogonal directions along which the mechanical properties of the FDM material have been characterised, i.e., they have been selected as the principal directions of the material. The x, y, and z coordinates are the machine coordinates. The constitutive equation for an orthotropic material is shown in equation (1) (Jones 1999): 21 31 1 2 3 12 32 1 11 2 3 2 213 23 1 2 33 3 23 23 2313 13 12 12 13 12 1 0 0 0 1 0 0 0 1 0 0 0 1 0 0 0 0 0 1 0 0 0 0 0 1 0 0 0 0 0 v v E E E v v E E E v v E E E G G G                                                                       (1) where i is the normal strain, \u03b3ij is the shear strain, Ei is the Young\u2019s modulus along i-th axis, Vij is the Poisson\u2019s ratio that corresponds to transverse strains in the j-th direction when load is applied in the i-th direction, Gij is the shear modulus in the i-j plane, \u03c3i is the normal stress, \u03c4ij is the shear stress in the i-j plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure1-1.png",
        "caption": "Fig. 1. Three Examples of Pivoted Shoes",
        "texts": [
            " Although disk brakes are growing in popularity, internal expanding-shoe brakes only are discussed. Types of Brake Shoe. A significant trend in brake design is the increasing use of floating shoes instead of the conventional type with a fixed pivot. This development has been pioneered by the proprietary brake manufacturers, but has not yet been adopted for those heavy vehicles in Great Britain whose brakes are of the chassis manufacturer\u2019s own design. I t is therefore of current interest to review the comparative merits of the two classes of shoe, those with a fixed pivot being typified in Fig. 1 while types of floating shoe are shown in Figs. 2 and 3. The fundamental difference between the two is that the floating shoe can centre itself in the drum when the brake is applied, whereas with a fixed pivot, concentricity is difficult to b Girling non-servo. P indicates position of maximum rate of wear. a Typical \u2019bus and truck shoe. c Girling hydraulic. at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from a Huck. b Lockheed slotted. c Lockheedtank. obtain and easily upset",
            " (3) Pivoted shoe mounted on a separate carrier. The first group might aptly be described as \u201cpivoted\u201d; a word which would then have to be avoided in relation to the floating types. \u201cNon-floating\u201d is not an attractive title, although adequate. \u201cAnchored\u201d is a possible alternative, but as ships float when at anchor it would be a misnomer if applied to shoes which do not float. \u201cRigid\u201d has been used, but is unsuitable since the shoe may deliberately be given flexibility. The following nomenclature is therefore proposed :- (1) Pivoted shop (Fig. 1). (2) Floating shoes : (a) Sliding (Fig. 2). (b) Articulated (Fig. 3). Other items are defined as follows :- Leading shoes: shoes in which the movement of the Trailing shoes: shoes in which the movement of the * An alphabetical list of references is given in Appendix 11, p. 51. drum over the lining is towards the pivot or abutment. drum over the lining is towards the applied load. Toe and heel: the ends of the lining in relation to the \u201cankle\u201d at the pivot or abutment. Shoe tip : the part of the shoe at which the operating load is applied; that is, the end remote from the pivot or abutment"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.25-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.25-1.png",
        "caption": "FIGURE 6.25",
        "texts": [
            " In this case, a relatively simple formulation can capture the increase in, for example, drag coefficient with aerodynamic yaw angle: ( ) GC A\u03b2CC\u03c1V 2 1F D\u03b2D0 2 D \u22c5+ = \u00f06:13\u00de where CD0\u00bc drag coefficient at zero aerodynamic yaw angle CDb\u00bc drag coefficient sensitivity to aerodynamic yaw angle b\u00bc aerodynamic yaw angle (or body slip angle surrogate) This formulation avoids the need for a knowledge of changing aerodynamic frontal area with attitude changes, which can be hard to obtain, and can be calculated from a fairly ordinary set of wind tunnel results. For the position shown in Figure 6.25 it is clear that for anything other than straight line motion it is going to be necessary to model the forces as components in the body-centred axis system. If we consider the vehicle moving only in the xy plane then this is going to require at least the formulation of a longitudinal force, Fx, a lateral force, Fy and a yawing moment, Mz, all in a body-centred axis system, usually located at the mass centre. Wind tunnel testing or CFD analysis is able to yield coefficients for all six possible forces and moments acting on the body, referred back to the mass centre"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002311_imece2016-65745-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002311_imece2016-65745-Figure3-1.png",
        "caption": "FIGURE 3: MESHING OF P3G PROFILE",
        "texts": [
            " A symmetrical contact behavior was chosen to avoid interpenetration on either side with normal stiffness of 1 and stabilization damping factor of 0.2 to avoid rigid body motion. The coefficient of friction was taken as 0.18 as determined from the studies of Grossman [4]. The body was meshed with lower order hexahedron mesh (Solid 185) with contact elements of 0.35 mm. The total number of elements varied according to the fit, averaged approximately 210,000. The meshing for P3G and P4C shaft hub connections are shown in Fig. 3 and Fig. 4 respectively. iii) Torsional bending load The analysis was performed by applying torsional bending load to the polygonal shaft hub connection. Only half of the connection was modeled due to symmetry along the axis with hub at the center of the connection as shown in Fig. 5. A torque of 80 Nm and a corresponding bending load of 4600 N was considered for the analysis to emulate the load from a spur gear. Initially, a rotating bending load was applied at 10\u00b0 difference for 120\u00b0 for P3G and 90\u00b0 for P4C connection along with static torsion to find the critical loading position as shown in Figs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001149_2014-01-1797-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001149_2014-01-1797-Figure13-1.png",
        "caption": "Figure 13. Architecture of clutch-1",
        "texts": [
            " The original output level of 60 kW was already high thanks to the lithium-ion battery. Figure 12 shows the battery architecture. This battery has high heat transfer performance due to its laminated cells and can be charged and discharged stably and quickly due to its manganese-based cathode with a spinel structure. For the second generation of the hybrid system, the maximum output power was limited precisely according to the discharge duration in order to boost the short-term maximum output power. Figure 13 shows the architecture of clutch-1. Clutch-1 is a dry clutch that is normally engaged in the EV mode. The engine and the motor are decoupled by charging oil to the clutch-1 cylinder to push the diaphragm. In order to start the engine, the diaphragm is returned by draining oil from the clutch-1 cylinder to couple the engine and the motor. At that time, the engine speed is raised by applying motor torque to start the engine. That is why it was important to reduce the engine startup time by improving the clutch-1 response so as to enable the vehicle to accelerate faster"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003051_0954407016629517-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003051_0954407016629517-Figure3-1.png",
        "caption": "Figure 3. Locations of the measurement systems in the research vehicle27 according to (a) the side view, (b) the bird\u2019s eye view and (c) the distribution of the lateral velocities: 1, Correvit-L sensor; 2, Correvit-Q sensor (front); 3, Correvit-Q sensor (rear); 4, impulse transmitter for measuring the steeringwheel angle; 5, lateral accelerometer; 6, angular velocity sensors.",
        "texts": [
            " Moreover, together with the application of a single-track vehicle model, the presented methodology of determining the side-slip characteristics for both steady-state conditions and non-steady-state conditions is justified, in particular because it requires a small number of identified coefficients. Nonetheless, a very important test is described below. The measurements presented in the subsequent sections of the paper were carried out using a light Ford Transit truck, equipped with sensors and a data acquisition system.27 The locations of the measurement equipment in the vehicle are shown in Figure 3. The measurement systems consist of the following: at University College London on June 5, 2016pid.sagepub.comDownloaded from (a) a Corrsys Datron Correvit-L opitical sensor for measuring the longitudinal displacement of the vehicle; (b) two Corrsys Datron Correvit-Q optical sensors for measuring the lateral displacement of the vehicle; (c) an impulse transmitter for measuring the steeringwheel angle; (d) a Lucas Schaevitz sensor for measuring the lateral acceleration of the vehicle; (e) a Murata piezo-gyroscope sensor for measuring the yaw velocity of the vehicle",
            " The following signals are measured: (a) the steering-wheel angle dk (impulse transmitter); (b) the longitudinal velocity vx (Correvit-L sensor); (c) the lateral velocities vQ1 and vQ2 at two points of the vehicle (Correvit-Q sensors); (d) the lateral acceleration ay (lateral accelerometer); (e) the yaw velocity _c (yaw velocity sensor). The lateral acceleration sensor and the yaw velocity sensor were mounted near the centre of gravity of the vehicle. The sampling frequency of the measured signals was 50 Hz. Selection of the gain of the measurement channels was preceded by many experiments for different levels of lateral dynamics. Figure 3 shows the locations of the measurement sensors and the variation in the lateral velocity across the longitudinal axis of the vehicle. On the basis of the lateral velocities vQ1 and vQ2 which are measured using CorrevitQ sensors, the yaw velocity can be determined from _c = v1y v2y l1 + l2 \u00f09\u00de and thus the derivative, namely the yaw acceleration \u20acc, can be obtained. The yaw velocity was also measured using a piezogyroscope Murata sensor. The results obtained in both directions were very close to each other and comparable, as can be seen in Figure 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000900_s10846-014-0157-z-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000900_s10846-014-0157-z-Figure2-1.png",
        "caption": "Fig. 2 Interpretation of angle \u03b1",
        "texts": [
            " When a stabilization scheme based on potential functions is used, conventional repulsive forces act over large repulsive distances. Generally, the direction of the repulsive force is normal to the obstacle, which greatly deviates the trajectory of the vehicle from its goal. Although conventional repulsive functions can be effective to avoid collisions, the trajectories of the vehicles worsens. In order to reduce these disadvantages we will use a function whose magnitude varies as a function of the angle \u03b1 between the vector of attractive force of the ith vehicle and the line from the vehicle to the obstacle, as shown in Fig. 2. Because the vehicle control is based on nested saturation functions, the magnitude of the repulsive function could be saturated too. Let Mri,o denote the repulsive force magnitude Mri,o = \u03c3br ( br (1 + cos(\u03b1))2 (1 \u2212 cos(\u03b1))2 [ 1 Lio \u2212 1 Lrep ]2 ) (33) where br is the bound of the magnitude of Mri,o , and could be chosen br > \u2016F at t,i\u2016, \u03b1 is an angle defined in Fig. 2, Lio is the current distance between vehicle i and the obstacle, and Lrep is the repulsive distance defined as follows Lrep = max{L0, 1 2 Lmin} where Lmin is the minimum repulsive distance chosen arbitrarily and L0 is a distance defined as a function of velocity of the vehicle L0 = Lxmax ( \u2016\u03be\u0307 i\u2016 x\u0307max ) + rX4 + Ladd (34) with Lxmax defined in Eq. 32, rX4 is the radius of the vehicle (distance from the center to the tip of any propeller), Ladd is an additional distance to be more conservative in the repulsive distance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002676_b978-0-12-409547-2.11182-5-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002676_b978-0-12-409547-2.11182-5-Figure4-1.png",
        "caption": "Fig. 4 Sample geometry and typical cell for fluorescence measurements.",
        "texts": [
            " Sample Illumination The most common arrangement uses the 90 geometry depicted in Fig. 3 rather than the 180 geometry common to absorption spectrophotometers. This geometry is suitable for weakly absorbing solution or diluted samples. For solids (samples adsorbed on solid surfaces such as polymers, paper, etc.) and for solutions that absorb strongly at the excitation wavelength, a front surface geometry, using triangular cuvettes or square ones at 30\u201360 respect to the incident light beam is preferable (see Fig. 4). Fluorescence is viewed in these cases from the face of the sample on which the exciting light impinges. In the front-face arrangement there is a risk that reflected light from the surface enters the emission monochromator resulting in large amount of stray light, particularly at an angle of 45 . If the solid sample is transparent, back-face illumination is also possible. In practice, most fluorescence measurements are taken in solution and 1 cm glass or fused silica cuvettes with four polished windows are used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001874_jbbbe.24.97-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001874_jbbbe.24.97-Figure1-1.png",
        "caption": "Figure 1: First configuration of IPMC actuated mechanism for fin deflection",
        "texts": [
            "224, University of Michigan Library, Media Union Library, Ann Arbor, USA-21/07/15,20:37:38) different voltages applied. Same actuation responses of the actuator are considered for analysis of both configurations. Deflection analysis is performed in Pro / Mechanism [14], an advanced simulation tool used to analyse kinematics of linked mechanisms. Performances of the two configurations are accessed for fin deflection by applying a new approach of virtual prototyping through computer aided simulations before going into manufacturing stage. Two proposed configurations of fin actuation mechanism are shown in Fig. 1 and 2. In first configuration, IPMC actuators are directly applied to the fin. Actuators act in cantilever form, fixed to base at one end and free to deflect at other end attached to the fin. Downward actuation of the actuator results in negative angle of attack of the fin and upward actuation results in positive. Whereas in second configuration, fin is actuated by the actuator through a connecting link (Link 2). Detailed description of the configuration with actuation principle is presented in [1]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001446_j.proeng.2014.06.045-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001446_j.proeng.2014.06.045-Figure3-1.png",
        "caption": "Fig. 3. Windsurf ergometer with (a) position sensor; (b) force sensor; (c) IMU on the mast foot; (d) heart rate monitor watch; (e) IMU on the athlete fore-arm; (f) embedded system attached to the athlete waist; (g) conditioning circuit of force sensor; (h) data acquisition and processing unit; (i) Bluetooth transmission unit; (j) ATmega328 microcontroller; (k) Li-ion battery.",
        "texts": [
            " along jb and the force Fa transmitted by the athlete to the rig is along j fa . F and Fa can be expressed by F = F.jb and Fa = Fa.jfa (7) F is the projection of Fa on the axis corresponding to jb and F is related to Fa by F = Fa. cos((F,Fa)) = Fa jb. jfa (8) jb = (S m S m C m \u2013 C m S m ) is + (S m S m S m + C m C m ) js + S m C m ks (9) jfa = (S fa S fa C fa \u2013 C fa S fa ) is + (S fa S fa S fa + C fa C fa ) js + S fa C fa ks (10) From relations (6), (8)-(10) the work can be deduced. 3. Material and methods Windsurf ergometer shown in Fig. 3 is composed of a 7.8 m\u00b2 One Design rig from BIC Sport company mounted on a wooden board with four wheels in order to calibrate the device in-door. A tape measure was attached to the boom to determine the location of the two athlete\u2019s hands on one hand, and a potentiometric position sensor (a) was also placed on the boom near the mast to determine the location of the windsurfer\u2019s hand that is closest to the mast in order to deduce the location of the point of application of F on the other hand. A strain gauge force sensor (b) was attached to the boom close to the inhaul to measure the magnitude of the force F transmitted to the rig by the athlete"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001996_sii.2015.7404997-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001996_sii.2015.7404997-Figure16-1.png",
        "caption": "Fig. 16. Forces acting on each link and accelerations of each link at 3\u00a9 t = 3.01 [sec] are shown at (a) as for y-direction and (b) as for z-direction. Also, at 4\u00a9 t = 3.95 [sec] they are shown at (c) as for y-direction and (d) as for z-direction.",
        "texts": [
            " All external forces acting on each link at the time of t = 3.95 are 0f\u03021 = [0,\u22120.274,\u22120.007]T, 0f\u03022 = [0,\u22120.007,\u22120.036]T, 0f\u03023 = [0,\u22120.841,\u22120.055]T. Therefore we got the same value of the acceleration obtained from 0f\u0302i/mi, so there is no contradiction between the relationship of acting force and acceleration. And Fig.14 shows the acting force in y- direction is maximum in the negative direction at the time of 4\u00a9. We can know that the shape of manipulator in this case has a shape close to a specific shape as shown in Fig.13 and Fig.16(c), (d). It shows that the force acting on each link will increase at the specific shape nearby. In this paper, we proposed the iterative calculation method to represent the constraint motion by using inverse dynamics calculation of the Newton-Euler method, and we have shown a solving method of the forward dynamics calculation. The results of simulation show the hand restraint motion of ma- nipulator can be expressed by proposed calculation method, and it is possible to calculate the force acting on each link"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003439_robio.2016.7866382-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003439_robio.2016.7866382-Figure2-1.png",
        "caption": "Fig. 2. The overall configuration of the robotic manipulation system.",
        "texts": [
            "00 \u00a92016 IEEE 562 by a HMM with the measured data, which is pre-processed by the multiresolution decomposition [17] because it can make the differences between physical constraints explicit. In the second part 2), the measured data is approximated to piecewise time polynomials, which are separated appropriately by the identification by a HMM. A sensory feedback experiment is shown for the effectiveness of the proposed method. The robotic manipulation system consists of left and right robot manipulators with multi-fingered robot hands, the configuration of which are illustrated in Fig. 2. In the following, the manipulators and multi-fingered robot hands are simply called as the arms and hands. The reference base frame \u03a3B is attached to the middle point on the robot base. The right arm has 6 joints qA\u22c6 \u2208 R6, where the script symbol \u22c6 \u2208 {r, l} stands for the right \u201cr\u201d and left \u201cl\u201d. The arm frame \u03a3A\u22c6 is attached to the end point, then the arm configuration is described by (BpA\u22c6 ,BRA\u22c6), where BpA\u22c6 \u2208 R3 and BRA\u22c6 \u2208 R3\u00d73 are the position vector and rotation matrix of \u03a3A\u22c6 with respect to \u03a3B respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002316_0959651815617883-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002316_0959651815617883-Figure10-1.png",
        "caption": "Figure 10. Schematic representation of the developed fourlink planar serial robot.",
        "texts": [
            " For the simulation part, a four-link planar serial robot with physical properties as the real robot, used in the experiment section, is considered. Figure 9 presents the developed robot. Table 3 indicates the leg length of at Middle East Technical Univ on December 31, 2015pii.sagepub.comDownloaded from each link of this robot and Table 4 indicates the angular limitations of all joints, which are considered in both simulation and practical sections. The parameters are shown in the schematic representation of the robot in Figure 10. First, it is assumed that one static obstacle exists in the workspace. The maximum relative velocity of all joints is vmax =0:007 m/s and the number of steps in each horizon is h=7. With the input data and obstacle position expressed in Table 5, the optimal path for the at Middle East Technical Univ on December 31, 2015pii.sagepub.comDownloaded from objective of minimum time is shown in Figure 11 which has the length of 1.02 m. The entire transitional time for this case is T=11 s and at each step it takes at most 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002748_ijmic.2016.075271-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002748_ijmic.2016.075271-Figure3-1.png",
        "caption": "Figure 3 Objective functions for region tracking",
        "texts": [
            " Here, the objective function describing the spherical attractive region as follows: ( ) ( ) ( ) ( )2 22 2 1 1 1 0d d df e x x y y z z r= \u2212 + \u2212 + \u2212 \u2212 \u2264 (7) where position error can be defined as 1 x d y d z d e x x e e y y e z z \u2212\u23a1 \u23a4 \u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5= = \u2212\u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u2212\u23a3 \u23a6 \u23a3 \u23a6 (8) If the region be considered as a circle in 2D or a sphere in 3D and the radius tends to zero, then the circular region or the spherical region become a point and it is the well-known set point regulation problem. Similarly, for spherical repulsive region, the objective function is written as ( ) ( ) ( ) ( )2 22 2 2 1 2 0d d df e x x y y z z r= \u2212 + \u2212 + \u2212 \u2212 \u2265 (9) with r1 and r2 are the radius of the respective desired regions. As shown in Figure 3, the desired region can be of annular shape depending on AUV\u2019s mission. Different shape and size of the region can be achieved using two or more objective function. The PE for the attractive region f1 and repulsive region f2 are considered to be ( ) ( ) ( ) 21 1 1 1 1 2 1 11 max 0, 2 if 0 2 0, otherwise p p K P e f K f f e \u23a1 \u23a4= \u23a3 \u23a6 \u23a7 >\u23aa= \u23a8 \u23aa\u23a9 (10) ( ) ( ) ( ) 22 2 1 2 2 2 2 12 max 0, 2 if 0 2 0, otherwise p p K P e f K f f e \u23a1 \u23a4= \u23a3 \u23a6 \u23a7 <\u23aa= \u23a8 \u23aa\u23a9 (11) Region tracking error 3 pe \u2208R can be calculated by partial differentiating the PE of the defined region in (10) ( ) ( )1 11 1 1 1 1 max 0, TT p p f ePe K f e e \u23a1 \u23a4\u2202\u2202\u23a1 \u23a4= = \u23a2 \u23a5\u23a2 \u23a5\u2202 \u2202\u23a3 \u23a6 \u23a3 \u23a6 (12) The region tracking error dynamics is given as 2 1 1 12 1 p Pe e He e \u239b \u239e\u2202= =\u239c \u239f\u2202\u239d \u23a0 (13) The analytical expression of the Hessian matrix in (13) is derived as 2 2 2 1 1 1 2 2 2 2 1 1 1 2 2 2 2 1 1 1 2 x x y x z x y y y z x z y z z P P P e e e e e P P PH e e e e e P P P e e e e e \u23a1 \u23a4\u2202 \u2202 \u2202 \u23a2 \u23a5\u2202 \u2202 \u2202\u23a2 \u23a5 \u23a2 \u23a5\u2202 \u2202 \u2202= \u23a2 \u23a5 \u2202 \u2202 \u2202\u23a2 \u23a5 \u23a2 \u23a5\u2202 \u2202 \u2202\u23a2 \u23a5 \u2202 \u2202 \u2202\u23a2 \u23a5\u23a3 \u23a6 (14) The Hessian matrix in terms of error is 2 1 2 1 1 2 1 4 2 4 4 4 4 2 4 4 4 4 2 x x y x z p x y y y z x z y z z e f e e e e H K e e e f e e e e e e e f \u23a1 \u23a4+ \u23a2 \u23a5= +\u23a2 \u23a5 \u23a2 \u23a5+\u23a3 \u23a6 (15) The total error vector is however 1 , TT T qe e e\u23a1 \u23a4= \u23a3 \u23a6 where the orientation error vector (eq) is given as d q d d e \u03b8 \u03b8 \u03c8 \u03c8 \u2212\u23a1 \u23a4 \u23a2 \u23a5= \u2212\u23a2 \u23a5 \u23a2 \u23a5\u2212\u23a3 \u23a6 \u03c6 \u03c6 (16) The objective is to develop a controller that will enable the dynamic system in (3) to track within the desired region given in (7) with the desired orientation while compensating for unknown restoring forces"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure18-1.png",
        "caption": "Fig. 18 Force components affecting bending moment in a femur, b tibia\u2014muscles considered",
        "texts": [
            " T1+ = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 Fhy1 x1 \u2265 0 Fmey1 x1 \u2265 I\u03011x L1 m1gy1 x1 \u2265 Cm1L1 Fm f y1 x1 \u2265 I\u03011bx L1 (49) T2+ = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 Fky2 x2 \u2265 0 Fmey2 x2 \u2265 I\u03012x L2 m2gy2 x2 \u2265 Cm2L2 Fm f y2 x2 \u2265 I\u03012bx L2 (50) The bending moment diagram has additional components in comparison to model A1, which are y components of muscle forces, and moments caused by x components of muscle forces due to the location of muscle insertion points outside neutral bending axis of the bones. The force components necessary to consider are depicted in Fig. 18. The muscle insertion points have been projected on the bone\u2019s neutral axis. The bending moment can also be presented in an incremental way, adding components according to their order along the bone. M1+ = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 \u2212Mh + Fhy1x1 x1 \u2265 0 Mme1 + Fmey1(x1 \u2212 I\u03011x L1) x1 \u2265 I\u03011x L1 m1gy1 (x1 \u2212 Cm1L1) x1 \u2265 Cm1L1 Mm f 1 + Fm f y1(x1 \u2212 I\u03011bx L1) x1 \u2265 I\u03011bx L1 (51) M2+ = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 Fky2x2 x2 \u2265 0 Mme2 + Fmey2(x2 \u2212 I\u03012x L2) x2 \u2265 I\u03012x L2 m2gy2(x2 \u2212 Cm2L2) x2 \u2265 Cm2L2 Mm f 2 + Fm f y2(x2 \u2212 I\u03012bx L2) x2 \u2265 I\u03012bx L2 (52) The evaluation of the expression should be done at each discrete point along the bone; however, only components for which the x condition is met should be added. For example, to evaluate bending moment at any point located between I1 and Cm1L1 on the femur in the same situation as in Fig. 18, the following formula should be used. M1(x1) = Mh + Fhy1x1 + Mme1 + Fmey1(x1 \u2212 I\u03011x L1) (53) In Eqs. 51 and 52, the moments at muscle insertions need to be computed according to the following formulas. Mme1 = I\u03011y L1 Fmex1 (54) Mme2 = I\u03012y L2 Fmex2 (55) Mm f 1 = I\u03011by L1 Fm f x1 (56) Mm f 2 = I\u03012by L2 Fm f x2 (57) The model can only take into account two muscle groups, thus individual muscle forces are not computed. The simulation is based on the principles of statics; therefore, dynamic loading is not taken into account, and the simulation is solved as a series of static cases; the rotation direction or velocity therefore does not affect the results in any way"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000123_b978-0-12-821614-9.00001-x-Figure1.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000123_b978-0-12-821614-9.00001-x-Figure1.2-1.png",
        "caption": "FIGURE 1.2-1",
        "texts": [
            " The application of Newton\u2019s laws of motion in their simplest mathematical form requires that the reference frame in which the motion is defined is an inertial reference frame (or inertial frame of reference). An inertial reference frame is a frame of reference in which the velocity of a mass point, with no applied external forces, is either zero or constant, i.e., there is no acceleration as measured relative to the reference frame. To illustrate the difference between inertial and non-inertial reference frames, we will compute the acceleration of a mass point undergoing circular motion about a point fixed in inertial space. Fig. 1.2-1A and B show a mass point, p, moving counterclockwise from position 1 to position 2 during a time intervalDt. In position 1 the location of the mass point is given bybr\u00f0t\u00de and at position 2 by br\u00f0t \u00feDt\u00de. In Fig. 1.2-1Awe have assumed a reference frame (Cartesian coordinate system) fixed in inertial space with the origin located at the center of the circular motion of the mass point. In Fig. 1.2-1B we specify a polar coordinate system to describe the motion (A) Position of point p defined in an inertial Cartesian coordinate system, with the orientation of unit vectors bex and bey parallel to the x and y-axes, respectively; (B) Position of the same point p as in (A) defined in a polar coordinate system, with the orientation of unit vectors ber\u00f0t\u00de and beq\u00f0t\u00de a function of q\u00f0t\u00de. of the mass point. The motion of the mass point in inertial space is not affected by the coordinate systems we use, but the mathematical description of this motion will be very different in the two coordinate systems. In Fig. 1.2-1A the location of mass point p, at time t (position 1), is given by br\u00f0t\u00de\u00bc x\u00f0t\u00debex \u00fe y\u00f0t\u00debey (1.2-1) where the superscript b designates a vector, and bex and bey are unit vectors parallel to the x and y-axes, respectively. When point pmoves from position 1 to position 2 over time Dt, the unit vectors bex and bey remain parallel to the x and y-axes. Hence, they are not a function of time. Differentiating br\u00f0t\u00de once with respect to time yields the velocity, d dt br\u00f0t\u00de\u00bc _br \u00f0t\u00de \u00bc _x\u00f0t\u00debex \u00fe _y\u00f0t\u00debey (1.2-2) and differentiating again yields the time rate of change of velocity, or acceleration, d dt _br \u00f0t\u00de\u00bc \u20acbr \u00f0t\u00de \u00bc \u20acx\u00f0t\u00debex \u00fe \u20acy\u00f0t\u00debey (1.2-3) In Fig. 1.2-1B the position of mass point p is described in the indicated polar coordinate system. q\u00f0t\u00de denotes the counterclockwise angle from the x-axis to vector br\u00f0t\u00de; hence, br\u00f0t\u00de\u00bc r\u00f0t\u00deber\u00f0t\u00de (1.2-4) where r\u00f0t\u00de is the distance at time t from the origin to point p, and ber\u00f0t\u00de is the unit vector that points from the origin to point p, i.e.,ber\u00f0t\u00de\u00bc br\u00f0t\u00de=jbr\u00f0t\u00dej (1.2-5) The unit vector beq\u00f0t\u00de is defined to be orthogonal (perpendicular) to ber\u00f0t\u00de, as shown. Unlike the rectangular coordinate system, unit vectors beq\u00f0t\u00de andber\u00f0t\u00demust be functions of time since their orientations relative to an inertial reference frame, in this case the Cartesian coordinate system x-y axes, change as point p moves. This can be seen in Fig. 1.2-1B, where these unit vectors do not remain parallel when point p moves from position 1 to position 2. Because these unit vectors are functions of time, computing the velocity and acceleration in polar coordinates is more involved than if we were in a Cartesian coordinate system. To obtain the velocity in polar coordinates, we differentiate Eq. (1.2-4) with respect to time, d dt br\u00f0t\u00de\u00bc _br \u00f0t\u00de \u00bc _r\u00f0t\u00deber\u00f0t\u00de \u00fe r\u00f0t\u00de _be r\u00f0t\u00de (1.2-6) Having the time derivative of a unit vector is not very convenient"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001753_ijmic.2015.072641-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001753_ijmic.2015.072641-Figure2-1.png",
        "caption": "Figure 2 Skyhunter (see online version for colours)",
        "texts": [
            " Section 5 is the results of the aircraft modelling identified and validation results are demonstrated as well. Section 6 gives conclusions of the paper. This paper gives an integrated and realisable experimental method for system identification. The aircraft used in the experiment is Skyhunter, which is commercially available to be employed as a flight test vehicle. Skyhunter is a low cost fixed-wing twin-boom aircraft, with electric drive. It has a wingspan of 1.78 m and a total weight of 2,323.4 g including data collection equipment inside. The frame of the aircraft is show in Figure 2. Physical properties of Skyhunter are shown in Table 1. Dynamics of the aircraft can be described by rigid body equations according to Newton\u2019s second law (Cook, 2012). A nonlinear model can be derived to express external forces and moments caused by gravity, propulsion and aerodynamics. The dynamics model is built in aircraft body axis. To describe the model, X-Y-Z body axis velocities are denoted u, v, w; X-Y-Z body axis angular rates are denoted p, q, r; body axis aerodynamic force are denoted X, Y, Z and the corresponding moments are denoted L, M, N"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000388_9780857094537.8.479-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000388_9780857094537.8.479-Figure1-1.png",
        "caption": "Figure 1. Schematic representation; (a) Front view of the rotor-bearing; (b) Perspective view of the pad.",
        "texts": [
            " \u2202 \u2202x h3 \u03bc \u2202P \u2202x + \u2202 \u2202z h3 \u03bc \u2202P \u2202z =6\u00b7\u03c9\u00b7RS\u00b7 \u2202h \u2202x +12\u00b7 \u2202h \u2202t (1) Where P(x,z) is the pressure distribution in the oil film, x and z are the rectangular co-ordinates, \u03bc is the absolute viscosity, RS is the radius of the shaft, h is the thickness of the oil film, \u03c9 is the rotational speed of the shaft and t is the time. The thickness of the oil film for each pad of the bearing can be obtained from geometrics parameters [7], eccentricity of the shaft and angular displacement of the pad as: h \u03b2 =RP-RS- sin \u03b2 \u00b7 yS+\u03b1\u00b7 RP+hP +cos \u03b2 \u00b7 xS+RP-RS-h0 (2) According to Figure 1, \u03b2 is the angular position in the pad, RP is the radius of the pad, hP is the thickness of the pad, h0 is the radial clearance, \u03b1 is the angular displacement of the pad, xS and yS are the position of the shaft in the local referential system. As previously mentioned, the pressure distribution is iteratively calculated with the temperature distribution. For this reason, the energy equation is applied to determine the temperature distribution in the oil film. In this case, no heat transfer was considered through the axial coordinate, because its effect is very small regarding the radial and circunferential directions in the system, as shown by Cameron [8]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003857_0954406215589843-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003857_0954406215589843-Figure11-1.png",
        "caption": "Figure 11. Beam element in rotating frame.",
        "texts": [
            " Moreover, the conversion from pressure in the air film to forces on the pad remains similar to the procedure of section Conversion of air film pressure to forces. It only requires calculation of the forces and moment with respect to the pad centre point instead of the pivot point. Not only the pad model, but also the rotor model has been improved for the 3D simulations. Rotor flexibility is very important in the dynamics of a high speed rotor system. Hence, the rigid body model has been replaced with a linear elastic finite element beam model of the rotor in a rotating at UNIV OF CONNECTICUT on June 4, 2015pic.sagepub.comDownloaded from frame. Figure 11 is a schematic view of one beam element. It has the following properties: . The position and orientation are defined by two nodes, one at each end point; . The cross section does not have to circular, but is constant for the element, hence no taper; . The nodes are on the centroid of the beam cross section; . The element local frame eb has a fixed position and orientation relative to the rotating frame er; . The displacements and rotations relative to frame er are small. This implies that the rotation speed of er equals the rotation speed of the rotor: _\u2019 \u00bc ; "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003420_ecce.2016.7855092-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003420_ecce.2016.7855092-Figure9-1.png",
        "caption": "Fig. 9. Density distribution of the iron loss at 7200 rpm and 16 Nm.",
        "texts": [
            " The fundamental component is almost the same with that of model A although the waveform is similar to triangular wave. Model D has both the effect of reducing the harmonic component by the disproportional gap and the effect of concentrating the field magnetic flux of PMs on the d-axis by the large flux barrier. Consequently, the harmonic component can be reduced without decreasing the fundamental component. Torque Area. Performance as the motor at high speed and high torque area was evaluated in this section. Fig. 9 shows density distribution of iron loss at 7200 rpm and 16 Nm. In model A, big iron loss caused by the q-axis magnetic flux occurs because the iron loss density in the rotor is distributed asymmetrically to the d-axis. In other models, the iron loss density distributed asymmetrically to the d-axis in the rotor is reduced significantly compared to model A. Moreover, the iron loss density in the stator of model B and D also decreases because of the disproportional gap. Fig. 10 shows the loss configuration at 7200 rpm and 16 Nm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003835_s11465-014-0304-z-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003835_s11465-014-0304-z-Figure9-1.png",
        "caption": "Fig. 9 Path generated for single robot by varying the parameters of IWO algorithm",
        "texts": [
            " The simulation experiments are conducted using MATLAB R2008 processing under Windows XP. All simulation results are applied on PC with Intel core Duo processor running 3.0 GHz, 4 GB RAM and a Hard disk of 160 GB. The simulation space is 40 cm by 40 cm rectangular environment with Gaussian obstacle functions at different locations as shown in Figs. 4\u20138. The safe range from robot to obstacles is 3 cm. The best parametric values of proposed algorithm have been selected after performing series of simulation experiments (shown in Fig. 9) on partially or totally unknown environments. It has been observed that some parameters are almost insensitive to select them, whereas other parametric values affect the performance of the proposed navigation system. Details of parameter values selection in IWO algorithm for mobile robot navigation are given in Table 1. The obstacle avoidance behavior and target seeking behavior has shown in Fig. 4. It is assumed that initially the robot is in an environment without prior knowledge about the location of obstacles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002464_978-981-10-1109-2_5-Figure5.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002464_978-981-10-1109-2_5-Figure5.8-1.png",
        "caption": "Fig. 5.8 Folding of CTR: 1Deployed position. 2\u20133 Intermediate positions. 4 Fully stowed position; VRRCTR = 0.33",
        "texts": [
            " Next such folded panels \u201ccollapse\u201d inside the unit. As Fig. 5.7 indicates, the IO system is substantially more complicated. However, the linear connections are compressed, which is advantageous for super-pressurized structures such as space stations or habitats. Moreover, since the side panels \u201ccollapse\u201d inside the unit, such scheme has better \u201cpacking\u201d capability. Alternatively, the folding of dPZ can be based on collapsing rigid concentric toric rings (CTR). The advantage of this concept is the lack of hinges, as shown in Fig. 5.8. As Fig. 5.8 indicates, CTR is a hingeless system. However, reaching full rigidity, sliding mechanisms, and sealing are major challenges. The same folding scheme, however, based on inflatability could avoid the last two problems relatively easily. Nevertheless, reaching proper rigidity of an elongatedmodular construction (without internal reinforcement) seems very difficult, if possible at all. Nonetheless, CTR system is neutral to under- and super-pressure. A CTR structure composed of a relatively few inflatable units seems rather feasible. Packing capability of CTR is also good (see Fig. 5.8). A low-fidelity six-unit octagonal dPZ has been fabricated. Similarly to the physical model of the hexagonal fPZM (see Fig. 5.1), the panels have been made of thick corrugated cardboard. For easy identification, the fPZMs have been made in contrasting colors. The units are connected by internal tubular elements. This connection gives one degree of freedom (1DOF), (rotation) between every two adjacent units. Each fPZM has two discrete states: stowed and erected. The transitions between these states are done manually"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000528_icra.2015.7140025-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000528_icra.2015.7140025-Figure5-1.png",
        "caption": "Fig. 5 Walk path of the experiment",
        "texts": [
            " In addition, the user is informed about an emergency switch position for stopping the cane robot, in a case of cane robot failure. Voice announcement is thus introduced to the cane robot for easy understanding of the robot state. The cane robot announces its movement and recommends behavior of the user based on the walk phase of the user. The experiments are done in two cases. In all cases, the user walks to a goal position following lines which are 1.5 [m] right and 1.5 [m] forward from a start position as shown in Fig. 5. In Fig. 5, the absolute coordinate is represented by vectors in the form [X,Y ], and the coordinate of the cane robot is represented by vectors in the form [x, y]. During experiment, the user walks normally, and the walking velocity of the user is about 0.5 [m/s]. In case 1, the users walks without tandem stance prevention method. In case 2, tandem stance prevention method is applied to the cane robot system. In the experiments, values of the parameters in the equations described above are empirically set as represented on Table II"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure12-1.png",
        "caption": "Fig. 12 Lower leg force diagram for model with muscles\u2014Grayed forces are not used in lower leg equilibrium",
        "texts": [
            "2 Model B: Muscle Actuation of the Knee Modeling a joint with muscles requires the separate derivation of knee forces. In this model, the knee joint is regarded as a pin joint, thus no internal knee moment exists. However, to restrain the lower leg from swinging the role of knee moment is taken by the muscle forces. Depending on the moment sign in the muscleless model, either the knee extensors or flexors are active. Therefore, to solve for the unknown muscle forces, the moment computed earlier can be used. The force diagram for the model with muscles is illustrated in Fig. 12. In the presented model, each muscle group can be regarded as a two-segment rope, which is represented in the Fig. 12 as a dotted line. If the muscle segments were virtually cut, the internal muscle forces can be applied to wrapping points and muscle insertion points. Point K then represents the knee cap, while point K \u2032 the knee pit. Points K and K \u2032 are muscle wrapping points and are considered to be rigidly attached to the upper leg at its end. Distances d and d1, depicted in Fig. 10b, represent distances between points P2 and K , and P2 and K \u2032 respectively. The moment equilibrium equation for the lower leg segment can be derived as follows"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure10-1.png",
        "caption": "Fig. 10. Magnetic field components of the permanent magnet motor: (a) stator iron field lines, (b) stator current field lines, (c) rotor iron field lines, (d) rotor magnet\u2019s field lines. The range of the contour lines is in brackets in Wb/m.",
        "texts": [],
        "surrounding_texts": [
            "linear system (24)] took less than 4% of the solution time of the276 motor magnetic field [assembly and solution of the nonlinear277 DAE (20)].278\nThe difference between B and \u2211\nk Bk should theoretically be279 zero. It was computed for FEM approximations, and the relative280 two-norm281 \u221a \u221a \u221a \u221a \u222b \u03a9 \u2016B \u2212 \u2211 k Bk\u20162 da\n\u222b \u03a9 \u2016B\u20162 da\n(27)\nremained below 5.3e-10, and the difference between the com-282\nputed A and \u2211\nk Ak remained below 7.5e-12 Wb/m for the283 studied time span.284 The stator field components are necessary for calculating the285 torque exerted on the rotor. In that case, applying (11) to the286 rotor iron and bars results in a total of four torque components:287 stator iron to rotor iron, stator iron to rotor current, stator current288 to rotor iron, and stator current to rotor current.289 Fig. 6 presents the total torque on the rotor computed with290 the MST and the sum of all torque components on the rotor.291 In this example, there is approximately a 1.5% difference be-292 tween the computed total rms torques. We have observed that293 the difference decreases with a refined mesh. This is consistent294 with the results presented in [20], where forces obtained with295 the MST and ESM (using external field) are compared with dif-296 ferent mesh densities (the sum of stator field components is the297 external field for the rotor).298 It is evident from Fig. 7 that a major part of the torque ripple299 produced by the squirrel-cage induction motor is due to stator300 iron to rotor iron interaction. In this example, there is also a301\nnotable phase shift between the stator iron to rotor iron torque 302 component and the other torque components. 303\nB. Permanent Magnet Motor 304\nFig. 8 presents the model geometry of the permanent magnet 305 motor and Table II the essential motor parameters. The rotor has 306 four surface-mounted magnets with 1-T remanence flux density 307 and 1.05 \u03bc0 permeability. The BH-curve of the iron cores and 308 shaft are presented in Fig. 3. The iron cores are assumed non- 309 conducting, whereas the shaft and magnets have a conductivity 310 of 4.3e6 S/m and 6.7e5 S/m, respectively. 311\nThe permanent magnet motor fields are decomposed into four 312 components: the stator iron field, the stator coil current field, the 313 rotor iron field, and the rotor magnets field. The magnetic field 314 and its components are presented in Figs. 9 and 10, respectively. 315\nThe relative two-norm (27) was used to measure the differ- 316\nence between B and \u2211\nk Bk . This remained below 4.8e-10, and 317\nthe absolute difference between A and \u2211\nk Ak remained below 318\n7.9e-12 Wb/m for the studied time span. For comparison, in [8], 319 the FR method was applied to a permanent magnet motor, and 320 there was approximately a 5% error in the magnetic flux density 321 in the motor air gap. 322",
            "Fig. 11 compares the total torques exerted on the rotor calcu-323 lated with the MST and with field decomposition. There is ap-324 proximately a 0.2% difference in the computed total rms torques.325 Fig. 12 presents the torques exerted by the stator iron and sta-326 tor coil current on the rotor iron and rotor magnets, respectively.327 It is evident from Fig. 12 that most of the permanent magnet328 motor torque and torque ripple is produced by stator iron to rotor329 magnets interaction.330\nVI. DISCUSSION331\nIn the examples in Section V, the motor magnetic fields were332 decomposed according to iron cores, current-carrying conduc-333\ntors, and permanent magnets. It is possible to decompose the 334 magnetic field in many other ways as well. For example, one 335 could refine, if necessary, the field generated by the stator core 336 into the field components of stator teeth and the rest of the sta- 337 tor iron. Furthermore, the field caused by stator coils can be 338 divided into field components caused by separate phases. The 339 refinement of field components also results in a number of new 340 torque components. 341\nThe results give new quantitative information on the parts of 342 the motor that produce torque. Such insights are useful in, for 343 example, minimizing the torque ripple. For example, the effect 344 of changes made to the shape of the stator tooth, the perma- 345 nent magnet, the squirrel-cage bar, winding distribution, and 346 the feeding current waveform can be seen in the corresponding 347",
            "torque components. Observing the sensitivity of the torque com-348 ponents to these changes can help the designer make more349 sophisticated decisions in optimizing parameter variation; for350 example, she can reduce the dimension of the feasible set.351\nVII. CONCLUSION352\nThis paper has presented a systematic way to isolate and353 quantify electromagnetic forces between different parts of an354 electric motor. Our approach is based on decomposing the mag-355 netic field to so-called magnetic field components generated by356 the magnetization and current density of the different motor357 parts. The forces and torques exerted by these magnetic field358 components were then computed with the conventional ESM359 force and torque expressions.360 The proposed analysis was applied to and demonstrated with361 two typical electric motors: a squirrel-cage induction motor and362 a permanent magnet synchronous motor. These two examples363 confirm that the proposed method can be successfully applied to364 different types of electric motors with nonlinear materials and365 eddy currents.366 Because the distinct interacting parts of an electromechani-367 cal device can be chosen freely, the proposed method offers a368 versatile tool to isolate and analyze force interactions in other369 devices as well.370\nAPPENDIX A371 CIRCUIT COUPLING372\nLet Q denote the index set of stator coils. Then, a stator coil373 with an index q \u2208 Q has a cross section \u03a9q consisting of coil374\nsides \u03a9+ q1 , . . . ,\u03a9+ qn and \u03a9\u2212 q1 , . . . ,\u03a9\u2212 qn . The superscripts + and375 \u2212 denote the orientation of the coil side with respect to the axial376 direction.377 To deal with the number of the turns and the direction of the378 current in the coils, let the source current density Js in the q:th379 coil cross section be380\nJs = Iq\u03b2qez (28)\nwhere Iq is the current in the q:th coil conductor, and \u03b2q is an 381 auxiliary function 382\n\u03b2q (p) =\n\u23a7 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8\n\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9\n+ Nqi\nSqi\nif p \u2208 \u03a9+ qi\n\u2212Nqi\nSqi\nif p \u2208 \u03a9\u2212 qi\n0 if p /\u2208 \u03a9q\n(29)\nwhere Nqi is the number of turns in the i:th coil side \u03a9qi , and 383 Sqi is the cross-sectional area of the i:th coil side \u03a9qi . 384\nThe voltage Vq and current Iq of the q:th coil are related by 385 equation 386\nVq = \u2202t\u03c8q + RqIq (30)\nwhere Rq is the series resistance of the coil and \u03c8q is the flux 387 linkage of the coil in the sense of 388\n\u03c8q = \u222b\n\u03a9q\nleA\u03b2q da (31)\nwhere le is the effective length of the motor. 389\nBecause this paper is focused on the decomposition of the 390 magnetic field and the computation of torque components, we 391 have chosen not to take into account the end effects of the motor 392 as external circuits. Coil end-winding and rotor end-ring circuit 393 models can be included if so desired; for example, see [21]. 394\nAPPENDIX B 395\nTABLE B1 CORE-IRON AND SHAFT-STEEL BH -PAIRS\nH (kA/m) B (T) H (kA/m) B (T)\n0.000 0.000 0.000 0.000 0.079 0.640 0.325 0.671 0.135 0.920 0.459 0.949 0.159 1.010 0.693 1.162 0.190 1.100 1.115 1.342 0.239 1.200 2.137 1.500 0.318 1.300 4.610 1.643 0.493 1.400 9.385 1.775 0.645 1.450 15.099 1.897 0.875 1.500 24.065 2.012 1.273 1.550 35.430 2.121 1.591 1.575 82.402 2.225 2.149 1.600 168.296 2.324 3.342 1.650 237.147 2.419 4.775 1.700 314.601 2.510 6.525 1.750 318.528 2.598 9.151 1.800 451.355 2.683 11.937 1.850 515.834 2.766 15.120 1.900 580.395 2.846 18.542 1.950 642.002 2.924 22.282 2.000 702.720 3.000 27.454 2.050 761.512 3.074 35.810 2.100 819.172 3.146 47.747 2.150 875.384 3.217 63.662 2.200 930.493 3.286 93.901 2.250 984.396 3.354\n(a) Core iron (b) Shaft steel"
        ]
    },
    {
        "image_filename": "designv11_64_0002933_gt2016-56900-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002933_gt2016-56900-Figure16-1.png",
        "caption": "Figure 16. Pressure field of oil film at different preload",
        "texts": [
            " Figure15 indicates the rotordynamic model of the rotor modeled with 55 elements, and the bearings locate at 5th and 35th node. Figure 15. Dynamic Model model of the compressor rotor Here, the load on both bearings is 4843N, and the load direction is on the pad (LOP). The comerical software DyRoBes_bperf was used to analyze the performance of the bearing. Considering the effect of bearing preload on the stability, the dynamic performance of bearing at different preloads from 0 to 0.5 were investigated. Figure 16 shows the pressure field of oil film at different preloads under rated operation speed (10700 r/min). It can be concluded that the oil film pressure will increase with the pad preload. Howerver, the two top pads will become unstable when the preload m is less than 0.2. Figure 17 shows the stiffness and damping coefficients of bearing at different preloads. It can be concluded that the stiffness increases and damping decreases with the increase of the preload. This type of tendency can cause a decrease in rotor 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002661_tdc.2016.7519956-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002661_tdc.2016.7519956-Figure8-1.png",
        "caption": "Fig. 8. Smooth conductors and pitted conductor behavior",
        "texts": [
            " During experimentation, it was noted that the C-Cor readings were more stable when the laser was aimed at the test conductor with a slightly horizontal angle rather than when the C-Cor was aimed just straight at it. This was due the behavior of the laser reflection on the conductor. Avoiding specular reflection increases the stability of the laser. This is why, during laboratory and field testing, the sample conductors are viewed at a horizontal angle between 15 and 30\u2070 to avoid specular reflection interferences (Fig. 8). Specular reflection and diffuse reflection while important for the stability of the readings are also now considered as a possible advantage in identifying pitting on the outside layer of aluminum strands. As shown in Fig. 8, pitted conductor and smooth conductor will produce different readings according to the angle of the C-Cor. From a wide angle to a small angle, the reading of a pitted conductor should not change due to the diffuse reflection which reflect the laser equally in every direction. However, from a wide angle to a small angle, the readings from a smooth conductor should increase as the smooth conductor acts like a mirror and the laser signals bounce back at angle equal to the incoming laser angle. This phenomenon could be used to identify the presence of aluminum strand pitting and with future development perhaps the severity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001238_978-90-481-9707-1_130-Figure12.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001238_978-90-481-9707-1_130-Figure12.2-1.png",
        "caption": "Fig. 12.2 Typical quadrotor",
        "texts": [
            " So the input is 6 degrees of freedom. If it needs to control 6 degrees of freedom, it has to have a minimum of 6 actuators. Or else the matrix is going to deficient. If only 4 actuators are provided as a quadrotor, it cannot possibly have independent control over 6 degrees of freedom. There are many hexrotors in the commercial world today. They all have six parallel thrust propellers spaced evenly around the circumference of a circle. All its thrusters are vertical, just like typical quadrotors as in Fig. 12.2. Therefore, they still result in rank-deficient matrices; in other words, there are no components from parallel thrusts that act in the plane perpendicular to the rotor axis. In this kind of configuration, these hexrotors work like quadrotors. Because all these thrusters are parallel and vertical, they can only provide linear force along Z axis, and torques around X, Y axes. Torque around Z is achieved indirectly through Coriolis forces resulting from differential angular velocities of the counterrotating propellers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000028_s0219455411004038-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000028_s0219455411004038-Figure1-1.png",
        "caption": "Fig. 1. The principal, local, and global coordinate systems of an element.",
        "texts": [
            " When the structure is in equilibrium in its deformed con\u00afguration for a given wrinkling status, the wrinkling convergence iteration cycle procedure is activated. A taut state is set up as default for the \u00afrst iteration of the \u00afrst load step, whereas the state of the elements at consecutive iterations is de\u00afned according to Eqs. (1) (3). The calculation of the stresses of an element in its principal directions requires the introduction of an additional coordinate system (1, 2, 3) in the direction of the principal directions as shown in Fig. 1. The coordinate systems described in Fig. 1 are the global (X, Y, Z), the local (xlocal, ylocal, zlocal\u00de, and the principal (1, 2, 3) coordinate systems. The global frame is used to de\u00afne the geometry, the applied loads, the structural sti\u00aeness matrix, and the global components of the nodal displacements (DOFs). The local frame is used to de\u00afne the element local sti\u00aeness matrices and stress vector. The principal frame is used to de\u00afne the elastic sti\u00aeness matrix and the stress vector of a wrinkled element. The orientation of the principal directions is calculated using the elementary second-order tensor transformation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000629_iccse.2015.7250378-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000629_iccse.2015.7250378-Figure7-1.png",
        "caption": "Fig. 7 The climbing robot teaching platform",
        "texts": [
            " Scheme 3: the integration of invention principle 1# and 5# Expanded structure of multi-functional folding drawing board is as shown in Figure 5. Through the analysis and comparison, we select the scheme 3 to solve this problem. We apply the teaching method to the students. In this mode, we use some platforms to ensure the teaching quality. For the students in different grades and majors, we apply different teaching platform. The educational robot teaching platform is as shown in Figure 6. The climbing robot teaching platform is as shown in Figure 7. Through problem based learning, the engineering undergraduate as the main innovation activities can greatly reflect the engineering undergraduate\u2019s dominant position, but also to maximize the engineering undergraduate\u2019s enthusiasm and initiative. This is an effective way to stimulate engineering undergraduate\u2019s practical innovative ability. This will be in favor of the improvement of engineering undergraduate\u2019s innovation thinking. Figure 8 is productions designed by engineering undergraduate using the TRIZ theory in problem based learning mode"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003666_s00022-015-0274-2-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003666_s00022-015-0274-2-Figure3-1.png",
        "caption": "Figure 3 a A regular tetrahedron abcd, b Pt for some 0 < t < 1 and c a flat folded state of the regular tetrahedron obtained by applying valley folds to dotted line segments and mountain folds to long dotted line segments",
        "texts": [
            " P is flattened explicitly by a continuous folding process of polyhedra {Pt : 0 \u2264 t \u2264 1} that satisfies the following: (i) the line segment Oh, where h is the midpoint of bc, is fixed on the x-axis, (ii) two faces abd and acd have no crease, (iii) there are points qt \u2208 Oh (0 \u2264 t \u2264 1) and rt \u2208 ah such that for each t, the face abc is divided into four triangles abrt, acrt, brth and crth, and bt(rt)\u2032h and ct(rt)\u2032h are attached to btqth and btqth, respectively, at, bt, ct and (rt)\u2032 denote the points on Pt corresponding to the points a, b, c and rt, respectively, (iv) the coordinates of at, bt, ct and qt are given by at = (6+2s \u221a 6+3s2\u221a 3(3+s2) , 0, 2(s\u2212\u221a 6+3s2) 3+s2 ), bt = ( \u221a 3, \u221a 1\u2212s2, s), ct = ( \u221a 3, \u2212\u221a 1\u2212s2, s), qt = ( \u221a 3(3+s2) 3+s \u221a 6+3s2 , 0, 0) where s = sin \u03c0 2 t (0 \u2264 t \u2264 1) (see Fig. 3), and (v) P0 = P and P1 is a flat folded state of P . A truncated polyhedron of a polyhedron P is the surface of a polyhedron obtained by cutting off vertices of P as solid polyhedra. A truncated regular tetrahedron Q has four regular hexagonal faces and four equilateral triangular faces. By Lemma 2.3, it is easy to see that the six hexagonal faces can be flattened continuously since they form a subset of the regular tetrahedron. We show how to flatten the four equilateral triangular faces in accordance to the continuous folding process for the rest of Q"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000165_ijvd.2019.105062-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000165_ijvd.2019.105062-Figure2-1.png",
        "caption": "Figure 2 7-DOF vehicle vibration model with dependent suspensions",
        "texts": [
            " The proposed multiparameter optimisation method is applied to improve mine car ride comfort, and the effectiveness of this method is validated by a road test. The mine car\u2019s front and rear suspension are both dependent. Using the features of the dependent suspension, a 7-DOF dynamics model is built to simulate the mine car. Assuming that the vehicle is symmetrical with the vertical XOZ plane, the centroid of the unsprung mass is in the XOZ plane, and the unsprung mass has only vertical and roll motion features; the vehicle body has vertical, roll, and pitch motion. Figure 2 shows the 7-DOF vehicle vibration model with dependent suspensions. The parameters in Figure 2 are defined as follows: body mass mb; unsprung mass of front and rear axle mf, mr; body pitch moment of inertia Ip; body roll moment of inertia I1; front axle unsprung mass pitch moment of inertia I2; rear axle unsprung mass pitch moment of inertia I3; left front, right front, left rear, and right rear suspension vertical stiffness ksa, ksb, ksc, ksd; left front, right front, left rear, and right rear suspension vertical damping csa, csb, csc, csd; left front, right front, left rear, and right rear tyre vertical stiffness kta, ktb, ktc, ktd; left front, right front, left rear, and right rear tyre vertical damping cta, ctb, ctc, ctd; horizontal distance from the centroid of the sprung mass to the front axle a; horizontal distance from the centroid of the sprung mass to the rear axle b; front wheel track 2Bf; rear wheel track 2Br; distance between the centre of two front suspension springs 2Lf; distance between the centre of two rear suspension springs 2Lr"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002454_1350650116652566-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002454_1350650116652566-Figure1-1.png",
        "caption": "Figure 1. Structure of spiral oil wedge sleeve bearing.",
        "texts": [
            " The motion equation, oil film thickness equation, and generalized Reynolds equation of spiral oil wedge sleeve bearing under dynamic loading are established, and the effect of dynamic loading and rotational speed on oil film cavitation and carrying capacity is studied by finite difference method and Euler method. Computation model of sleeve bearing based on dynamic loading conditions Oil film thickness equation of spiral oil wedge sleeve bearing The structure of spiral oil wedge sleeve bearing is shown in Figure 1. The bearing has three tilted spiral oil wedges in the circumferential direction and both ends of every oil wedge have oil feed holes 2 and oil return holes 1. The special structure of sleeve bearing makes the oil film thickness to be calculated in the cylindrical surface and eccentric circular surface.17 Figure 2 shows the oil film thickness calculation diagram under dynamic loading. O is the bearing center, O2 is the axis center that is denoted by (x, y), Oh are the calculation points of counter-clockwise rotation along x-axis, and is attitude angle, cos \u00bc x=e, sin \u00bc y=e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001782_iros.2015.7354084-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001782_iros.2015.7354084-Figure9-1.png",
        "caption": "Figure 9. Velocity-based detection mechanism.",
        "texts": [
            " One end of Linear Spring C is connected to Pin C1 attached to Plate C, and the other end is connected to Pin C2 in Frame C. One Guide Bar C is attached to each Pawl C. Each Guide Bar C is inserted into each Guide Hole C in Plate C. Gear D meshes with Gear A. A rotary damper is connected to Gear D. Pawl D is connected to the shaft of the rotary damper. One end of Linear Spring D is connected to Pawl D, and the other end is connected to Frame D. Frame D is mounted on Frame C. Switch C can cut off the electric power supply to all of the robot\u2019s motors. Switch C is pressed by Pin C when Plate C is rotated. Fig. 9 shows the mechanism which mechanically detects the unexpected robot motion on the basis of the angular velocity of Shaft C. The damping torque generated by the rotary damper and the spring torque generated by Linear Spring D act on Pawl D, when Gear D is rotated by Gear A. As the velocity of Gear A (i.e. Shaft C) increases, the damping torque increases. Pawl D rotates by the torque difference between the damping torque and the spring torque, and locks Plate A, if the velocity of Shaft C exceeds the detection velocity level. The detection velocity level is adjustable by changing the attachment position of Linear Spring D as shown in Fig. 9. We can change the attachment position of Linear Spring D by using the adjustment mechanism which we developed for adjusting the detection velocity level as shown in [12]. Fig. 10 shows the mechanism to mechanically lock Shaft C. After Plate A is locked, each Pawl B slides along each Guide Hole A in Plate A because Plate B is still being rotated by the motor (Fig. 10(b)). Then one of the three Pawl Bs is hooked onto the inner teeth and rotates Plate C (Fig. 10(c)). By Figure 7. Velocity-based safety device"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000466_jae-140095-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000466_jae-140095-Figure5-1.png",
        "caption": "Fig. 5. Scheme of the cantilever beam bonded with the PVDF film.",
        "texts": [
            " The parameters\u2019 value and Eqs (6) and (7) show that if the light intensity is increasing, the photocurrent will increase and the photoresistance will decrease, and therefore the maximum strain of the PVDF will increase and the illumination time constant will decrease. The PLZT/PVDF hybrid drive is mainly used for shape adjustment of space shell structure, in which the PVDF actuator is attached on the shell\u2019s surface. Now taking the cantilever beam bonded with a PVDF actuator as the study object, effect of the PLZT/PVDF hybrid drive is simulated via computing the end displacement of the cantilever beam. The cantilever beam bonded with the PVDF film is shown in Fig. 5, whose length, breadth, and thickness is L, b, and h, and the length, breadth, and thickness of the PVDF film is La, ba, and ha. Table 3 Parameters of the cantilever beam Length (L) Breadth (b) Thickness (h) Young\u2019s modulus (E) 0.12 m 0.008 m 0.0005 m 1.6 \u00d7 109 Pa Table 4 Parameters of the PVDF film Parameter Value Unit Length (La) 34 mm Breadth (ba) 8 mm Thickness (ha) 0.05 mm Young\u2019s modulus (Ea) 2500 MPa Piezoelectric coefficient (d31) 18 pC/N Relative permittivity (\u03b5r) 9 Resistance (Rf ) 3.7 \u00d7 109 \u03a9 When the PVDF actuator strains, force and moment on the cantilever beam are: Na = habaEaS[u(x)\u2212 u(x\u2212 La)] (15) Ma = h+ ha 2 \u00d7 habaEaS[u(x) \u2212 u(x\u2212 La)] (16) Where, Ea is the Young\u2019s modulus of the PVDF film; S is the strain of the PVDF film; u(x) is the unit step function"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002577_1350650116656982-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002577_1350650116656982-Figure3-1.png",
        "caption": "Figure 3. Scheme of the system for measuring the oil film resistance in the worm meshing: 1 \u2013 worm-wheel, 2 \u2013 dielectric separators (resocard), 3 \u2013 hub, K \u2013 mercury collector, Osc \u2013 oscilloscope, mV \u2013 millivoltmeter, V \u2013 voltmeter, R0, R1, R2 \u2013 resistors.",
        "texts": [
            "7 The theory together with the research3 indicates that the type of friction emerging in the meshing of a worm gear is influenced first of all by factors such as: \u2013 hydrodynamic speed, \u2013 oil viscosity, \u2013 carried load. The value of the oil film resistance between the worm and the worm gear has been adopted as the criterion of the existing type of friction in the meshing during the cooperation of these elements. The wormwheel was electrically isolated from the remaining elements of the reducer, and after applying DC voltage of 40mV to the worm and worm-wheel, the emergent oil film3 remained the only significant resistor in the system (Figure 3). The resistance of the oil film in a worm gearing was calculated twice each time for both directions of electric current, all this for the purpose of eliminating the influence of possible polarization and the average of the obtained values was taken. The results are presented in Figures 4 and 5. The changes in resistance of the oil film in the meshing of a worm gear with bronze wormwheel are presented in Sabiniak.3 Moreover, during the whole period of research an image of the type of friction was observed on the screen of oscilloscope"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001306_s1063454114040086-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001306_s1063454114040086-Figure4-1.png",
        "caption": "Fig. 4. Short time motion without jumps in the control force, T = T2, T2 = 0.5T1.",
        "texts": [
            "3), (2.4), (7.1) in this set ting may not be solved by minimizing functional (3.1) via the Pontryagin maximum principle, because in this case the number of arbitrary constants in the solution is insufficient. Contrariwise, a solution of a sim ilar extended boundary value problem using the generalized Gauss principle may be constructed; to this end it suffices to increase its order by two. The numerical results for the generalized boundary value prob lem with T = T2, T2 = 0.5T1 is represented in Fig. 4. The graph of the dimensionless control shows that it proved possible to eliminate the jumps in the control force at the terminal times of motion of the system. We point out that the application of the generalized Gauss principle to dampen the oscillations of the mechanical system in question is preferable. However, this method is rather restricted in scope. Numerical x0'' 0( ) x0'' T( ) 0.= = VESTNIK ST. PETERSBURG UNIVERSITY. MATHEMATICS Vol. 47 No. 4 2014 ON THE RELATIONSHIP BETWEEN CONTROL THEORY 187 analysis has shown that the solutions depend quite essentially on the parameter \u03bb = T/T1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure3.13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure3.13-1.png",
        "caption": "Fig. 3.13 Electric circuit with a variable shunt conductivity",
        "texts": [
            " Then, the subsequent voltage value VB2 L \u00bc VG L V C2 L m21 i VC2 L \u00f0m21 i 1\u00de\u00feVG L \u00bc 12 4:5 0:7272 4:5 \u00f00:7272 1\u00de\u00fe 12 \u00bc 3:6453: Case 3 Recalculation of the load voltage at the common change of the load RL and resistance r0N Let the regime change be given as C1 \u2192 C2 \u2192 B2. Common regime change (3.33) m21 \u00bc m21 i m21 L \u00bc 0:7272 0:6 \u00bc 0:4363: Resultant voltage value (3.34) for the point B2 VB2 L \u00bc VG L V C1 L m21 VC1 L \u00f0m21 1\u00de\u00feVG L \u00bc 12 6 0:4363 6 \u00f00:4363 1\u00de\u00fe 12 \u00bc 3:6453: 3.2 Circuit with a Series Variable Resistance 69 3.3 Circuit with a Shunt Variable Conductivity 3.3.1 Disadvantage of the Known Equivalent Circuit Let us consider an active two-pole circuit with a base load conductivity YL and variable auxiliary load or shunt regulating conductivity yN in Fig. 3.13. This circuit has a practical importance for a current regulation. At change of the load conductivity from the short circuit SC to open circuit OC for the specified shunt conductivity yN, a load straight line is given by expression (3.1) IL \u00bc YiV OC L YiVL \u00bc ISCL YiVL; \u00f03:35\u00de where VOC L is the OC voltage; the internal conductivity Yi and SC current ISCL are the parameters of the Norton equivalent circuit in Fig. 3.14. For our two-pole circuit we have 70 3 Generalized Equivalent Circuit of an Active Two-Pole \u2026 Let the conductivity yN varies from y1N to y2N ",
            " This brings up the problem of determination in the relative or normalized form of the conductivity value yN regarding of these characteristic values. Therefore, the obvious value yN = 0 is not characteristic ones concerning the load. y 0 N =\u221e V G L y V N y I N Y I i =0 Y 0 i y N Y V i =\u221e I G L Y i Y L1 0 V L G I L Let us view the load characteristic values. Both the traditional values YL = 0, YL = \u221e, and YG L will be the characteristic values according to Fig. 3.18. 3.3.3 Relative Operative Regimes. Recalculation of the Load Current Let us consider a common change of load YL and variable shunt conductivity yN of the circuit in Fig. 3.13. The corresponding load straight lines are shown in Fig. 3.20. In this figure, we get the two bunches with parameters YL and yN. Let an initial value of the variable element be y1N and subsequent value is y2N . Similarly, an initial value of the load equals Y1 L and subsequent one is Y2 L . A 2 A 1 F 2 F 1 B V B I B 2 B 1 C V C I C 2 C 1 Y L =0 Y G L Y L =\u221e y 2 N Y 2 L y 0 N y V N V G L y 1 N Y 1 L I G L y I N 0 V L G I L Fig. 3.20 Common change of load YL and variable conductivity yN 74 3 Generalized Equivalent Circuit of an Active Two-Pole \u2026 Case 1 Definition of the relative operating regime at the load change The cross ratio m1 L of the four point, three of these are the characteristic points CV, CI, G of the line y1N , and the fourth C1 is the point of the initial regime YL 1, has the view m1 L \u00bc \u00f0CV C1 CI G\u00de \u00bc C1 CV C1 G CI CV CI G \u00bc \u00f00 Y1 L 1 YG L \u00de \u00bc Y1 L Y1 L YG L : \u00f03:46\u00de The cross ratio for the points BV ; B1; BI ; G of the line y2N has the same value m1 L \u00bc \u00f0BV B1 BI G\u00de \u00bc B1 BV B1 G BI BV BI G : \u00f03:47\u00de The points CV, G (and BV, G) are the base points, and CI (and BI) is a unit one"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001742_s00542-015-2460-4-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001742_s00542-015-2460-4-Figure4-1.png",
        "caption": "Fig. 4 The model of the air bearing spring, showing the disk and slider",
        "texts": [
            " Figure 3 shows that the effect of the stiffness of the air bearing on the transient shock simulation. The stiffness of the air bearing was halved by altering the spring constants at the leading and trailing edges. The fact that the vertical 1 3 force for these two cases was almost identical shows that the effect of the stiffness of the air bearing over a wide range was negligible as in a previous study (Zeng and Bogy 2002). The slider was connected to the disk using four linear springs, as shown in Fig. 4. These springs were located at the center of trailing edge (TE), the leading edge (LE), inner edge (IE) and outer edge (OE) of slider. When an external shock was applied, it was transmitted through four bolt points. The large mass method is one of methods to describe the external shock transmission, whereby a large mass was connected to the four bolt points as shown in Fig. 5. And, the external shock is applied to the large mass to investigate the shock response. With the contact model, it is necessary to establish which surfaces are contact and target surfaces"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003003_978-3-319-42402-6_8-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003003_978-3-319-42402-6_8-Figure1-1.png",
        "caption": "Fig. 1 a Rotor model prepared in the Samcef Rotors software, b Bearing model",
        "texts": [
            " Experimental identification of stiffness and damping coefficients of an axial foil bearing was shown in the article [18]. 2 The Calculations of Bearing Dynamic Coefficients The numerical model of the rotor, which have been used for the sensitivity analysis of the method, was created on the basis of the test rig called SpectraQuest Machinery Fault and Rotor Dynamics Simulator. This test rig is located at the Szewalski Institute of Fluid-Flow Machinery PAS, in Gda\u0144sk. The numerical model was created in the Samcef Rotors software (Fig. 1). The model consists of a shaft rotating at 2800 rpm and bearings modeled using the stiffness and damping coefficients in the orthogonal and cross-coupling directions. The rotor during the simulation had been excited by a known value of excitation force, in the central part, in the X direction. Then the vibration amplitude of the rotor was measured in each bearing in X and Y directions. Then, the simulation was repeated, this time using the same excitation force, but model has been excited in the Y direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001587_aic.690060123-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001587_aic.690060123-Figure4-1.png",
        "caption": "Fig. 4. Dispersion about the average angle of reflection is normal in character.",
        "texts": [
            " When the data from different incident angles were assembled, it was found that the standard deviation from the angle of reflection appeared to vary with the incident angle. As shown in Figure 3, the standard deviation from the average angle of reflection varied from approximately 9 to 10 deg. for an incident angle of 50 deg. to 5 to 6 deg. for an incident angle of 80 deg. For incident angles greater than 70 deg., the mean reflected angle was approximately equal to the incident angle, and the distribution of angular deviations was normal in shape. A sample distribution is shown in Figure 4. Results from the bounce of a soft rubber cube on a smooth varnished wood surface and a hard Lucite cube against a smooth steel surface are also shown in Figure 3. Apparently friction plays an important role in setting the standard deviation. With rubber cubes the average angle of reflection was found to be lower than the angle of incidence for high angles of incidence and higher for low angles, the difference being more pronounced at high angles of incidence. With Lucite cubes the average angle of reflection was always more than the angle of incidence, the difference slightly increasing with decreasing angle of incidence"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003256_978-981-10-2502-0_9-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003256_978-981-10-2502-0_9-Figure5-1.png",
        "caption": "Fig. 5 Sketch of the center-of-mass position Ri and the orientational unit vector u\u0302i for the i th particle both for a rod-like and plate-like particle rod i",
        "texts": [
            " This demonstrates that DDFT is a reliable microscopic theory both if hydrodynamic interactions are included or ignored. Density functional theory can readily be extended to rod-like particles which possess an additional orientational degree of freedom described by a unit vector u\u0302 [3, 13, 28]. orientation vector center-of-mass coordinate \u00fb R i i A configuration of N particles is now fully specified by the set of positions of the center of masses and the corresponding orientations {Ri , u\u0302i , i = 1, . . . , N }, see Fig. 5. Examples for anisotropic particles include (1) molecular dipolar fluids (e.g. H2O molecule) (2) rod-like colloids (e.g. tobacco-mosaic viruses) (3) molecular fluids without dipole moment (apolar), (e.g. H2 molecule) (4) plate-like objects (clays) The canonical partition function for rod-like particles now reads [27] Z = 1 h6N N ! \u222b V d3R1 ... \u222b V d3RN \u222b R3 d3 p1 ... \u222b R3 d3 pN \u00d7 \u222b S2 d2u1 ... \u222b S2 d2uN \u222b R3 d3L1 ... \u222b R3 d3LN e\u2212\u03b2H (73) with the total Hamilton function H = N \u2211 i=1 { p2i 2m + 1 2 Li ( \u00af\u0304\u0398)\u22121Li } + 1 2 N \u2211 i, j=1 v(Ri \u2212 R j , u\u0302i , u\u0302 j ) + N \u2211 i=1 Vext(Ri , u\u0302i ) (74) which comprises the kinetic energy, the pair interaction energy and the external energy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002622_ssp.251.49-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002622_ssp.251.49-Figure2-1.png",
        "caption": "Fig. 2. Wire rope defects",
        "texts": [
            " All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (#69341303, Purdue University Libraries, West Lafayette, USA-31/08/16,14:15:30) The contact of wire rope with the pulleys, other ropes, metal surface with sharp edges causes various rope defects: \u2022 mechanical wear-and-tear; \u2022 breaking from fatigue of cable bending; \u2022 rope corrosion; \u2022 break due to rope tension overload (Fig.2. a); \u2022 twisted rope (Fig. 2. b); \u2022 break of rope strand; \u2022 loosening of one or several wires; \u2022 bending the rope (Fig. 2. c); \u2022 break of one or several wires (Fig.2. d); \u2022 other defects [2,3]. Different methods and practical equipment are used to detect and diagnose wire rope defects. Visual method. During visual inspection, the arrangement of individual parts, nodes, units, and elements are checked as well as the trajectory of motion, the condition and cleanliness of the surfaces, deformations, fractures, the position of fastenings and fittings [2\u20134]. The efficiency of the visual inspection method depends on the experience and the ability of the operator to concentrate on individual details",
            "52 \u2013 coefficient of 1st natural frequency, l \u2013 length of wire, E \u2013 Young\u2019s modulus, I \u2013 moment of inertia of wire, \u03c1 \u2013 specific weight of material, F \u2013 area of wire cross-section. Frequency of broken wire f: \u03c0 \u03c9 2 =f (2) Actual broken wire diameter: d=0.26mm = 0.26 \u00b7 10-3 m. Actual length of broken wire: l=17.0mm = 17.0 \u00b7 10-3 m. Desired natural frequency of actual wire is about 1400 Hz, but cantilever clamping of wire condition differs from theoretically defined one. For research of rope and broken wire dynamic behaviour there is designed and produced special test rig, presented in Figure 2. Frame of test rig weights about 250 kg and casted from cast iron. The holders 3 and 4 are firmly affixed to the analysed wire rope 2 on the stand 1. One end of the wire rope 2 is affixed to the holder 3 and extended horizontally by the tensioning screw in the holder 4 affixed to the other end of the wire rope. Wire rope tensile force is approximately 5kN. The electrodynamic mini vibrator 5 is firmly pressed against the analysed rope for using a rod; the electrodynamic mini vibrator 5 is affixed to stand 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure2.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure2.9-1.png",
        "caption": "Fig. 2.9 (a) Key parameters and (b) flow profile of a slit flow",
        "texts": [
            " Roughly, Re < 1000 indicates dominance of the viscous force leading to laminar flow; and Re > 1000 reflects dominance of the inertial force leading to turbulent flow. Considering that the characteristic length is smaller than a centimeter and the \u201ccharacteristic velocity\u201d is limited in a human body and in biomedical devices, corresponding flows are often laminar. In more extreme cases, inertial effects can be ignored as Stokes flow or creeping flow. For a liquid flow in a rectangular slit with length L, width W, and height H along the slit length direction driven by a fluidic pressure gradient along slit length as shown in Fig. 2.9a, further considering that L W H, the governing equation can be expressed as \u03bc d2u dy2 \u00bc dp dx \u00f02:30\u00de where u( y) is the flow velocity along the slit length, p is the fluidic pressure, x is the position along the slit length, and y is the position along the slit height. By letting the middle position in the slit be y\u00bc 0 such that u(H/2)\u00bc 0 and u( H/2)\u00bc 0 (Fig. 2.9b), according to the no-slip boundary conditions, we obtain u y\u00f0 \u00de \u00bc dp dx 1 \u03bc H2 4 y2 \u00f02:31\u00de The liquid should flow from a position with a higher pressure to a position with a lower pressure; therefore, the flow velocity u should have a direction along a negative pressure gradient dp/dx. The fluidic pressure decreases along the flow direction caused by friction and kinetic energy change. Hence, the fluidic pressure at 2.4 Fluidic Properties 47 the slit inlet (Pin) is higher than the pressure at the slit outlet (Pout), and the negative pressure gradient can be estimated as \u0394P/L, where the gauge pressure is \u0394P \u00bc Pin Pout"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure9-1.png",
        "caption": "Figure 9 A schematic of specimen configurations for three point bending (see online version for colours)",
        "texts": [
            " The test was conducted on an Instron 4411 load frame which is equipped with a built-in 4448.22 N (1000 lb) load cell. The test was displacement control where the speed of crosshead was kept at 1.27 mm/min (0.05 in./min). Toe compensation was carried out after test. Figure 8 depicts a schematic of the three-point bending fixture. The specimen is a rectangular bar, with the length of 76.2 mm (3 in), width of 25.4 mm (1 in) and thickness of 7.62 mm (0.30 in). Four different specimen configurations as shown in Figure 9 were designed in order to fully verify the validity of constitutive model derived above. The test matrix is shown in Table 3. Each sample had four replicas for three point bending tests. Load vs. deflection curves were plotted for different configurations as shown in Figure 10. It is observed that the slope of 3P3 is highest, followed by 3P2 and 3P1, and that of 3P4 is lowest. The deflection of all the parts at 200N was picked as the point of comparison among experimental, numerical, and analytical results"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001598_1.g000127-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001598_1.g000127-Figure2-1.png",
        "caption": "Fig. 2 Steering problem in the presence of wind with an uncertain spatiotemporal velocity field for a controlled object whose motion is described by Eq. (15).",
        "texts": [
            " It will be referred to as the underactuated kinematic model, which is described by the following equations [19,30]: _\u03be \u03bd \u03c9 t;\u03be \u0394\u03c9 t;\u03be ; _\u03bd \u03d6\u00d7\u03bd; \u03be 0 \u03be; \u03bd 0 \u03bd (15) where \u03d6 is the new control input and \u00d7 denotes the cross-product operation. Note that _\u03bd \u03d6 \u00d7 \u03bd is, by definition, perpendicular to the air velocity \u03bd at all times, and thus it can be written as follows _\u03bd \u03d6 \u00d7 \u03bd _\u03bd\u22a5\u03bd _\u03bd\u22a5\u22a5\u03bd , where _\u03bd\u22a5\u03bd and _\u03bd\u22a5\u22a5\u03bd are two mutually orthogonal components of _\u03bd, which span a plane that is perpendicular to the air velocity \u03bd. The situation is illustrated in Fig. 2 (note that _\u03bd\u22a5\u22a5\u03bd is pointing into the page). Note that only the components of\u03d6 that are perpendicular to \u03bd can affect the motion of the underactuated CO. It is also interesting to highlight that the state constraint j\u03bd t j 1, for all t \u2208 0; tf (hard constraint), is now encoded in the new equations of motion of the CO given in Eq. (15). In particular, _\u03bd is now perpendicular to \u03bd, and thus the CO is constrained to travel with constant airspeed (there is no component of _\u03bd parallel to \u03bd to change the CO\u2019s airspeed)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002975_aim.2016.7576813-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002975_aim.2016.7576813-Figure2-1.png",
        "caption": "Fig. 2. (a) Hall model and (b) basic personal space",
        "texts": [
            " We define the state of the robot as r = (xr,yr,\u03b8r,vr), where the position is (xr,yr), the orientation is \u03b8r, and the linear velocity is vr. Figure 1 illustrates human states including human position, orientation, velocity, and field of view, robot states including the robot position, orientation, and velocity, and the relative localization between a person and a robot. B. Personal space The concept of personal space, first proposed by Hall [9], divides the space around a human into four main zones, namely intimate zone, personal zone, social zone and public zone in terms of the range from comfort to social activity, as shown in Fig. 2(a). In order to model the space around a person, in this paper, we use of a two-dimensional Gaussian function. The basic personal space surrounding the person pi is defined as a function f p i (x,y) that has its maximum at the center (xp i ,y p i ) and gradually descends away from (xp i ,y p i ). Equation (1) represents the model of the basic personal space around a person pi. f p i (x,y) = Ap exp ( \u2212 ((dcos(\u03b8 \u2212\u03b8 p i )\u221a 2\u03c3 px 0 )2 + (dsin(\u03b8 \u2212\u03b8 p i )\u221a 2\u03c3 py 0 )2 )) (1) F = {Fx \u2208 Fc : (|S|> |C|) \u2229 (minPixels < |S|< maxPixels) \u2229 (|Sconected|> |S|\u2212 \u03b5)} (2) where d and \u03b8 are computed as follows: d = \u221a (x\u2212 xp i ) 2 +(y\u2212 yp i ) 2, (3) \u03b8 = atan2((y\u2212 yp i ),(x\u2212 xp i )), (4) where (x,y) is the coordinate of a point around the person pi, (x p i ,y p i ) and \u03b8 p i are the position, orientation of the person pi, respectively, Ap is the selected amplitude, and \u03c3 px 0 and \u03c3 py 0 are the standard deviations of the Gaussian function. A set of three primary parameters used for the function f p i (x,y) is [Ap,\u03c3 px 0 ,\u03c3 py 0 ]. Fig. 2(b) shows an example of the basic personal space in contours with the parameter set [1,0.45,0.45]. C. Hybrid Reciprocal Velocity Obstacles Model (HRVO) The HRVO, introduced by Snape et al. [16], is an extension of the RVO [15]. This technique is a velocity-based approach [14] synthesizing the motion of neighbourhood agents for collision avoidance. A construction of the HRVO of a robot and a human is illustrated in Fig. 3 and detailed description of the method can be found in [16]. We suppose that a set of human P and a set of dynamic and static obstacles O appear in the robot\u2019s vicinity",
            " The system is detailed in the following sections. A. Approaching Pose Estimation 1) Extended Personal Space: As presented in Eq.(1), the size and shape of the basic personal space depend on the primary parameters [Ap,\u03c3 px i ,\u03c3 py i ]. In this study, we incorporate the human states including human position, orientation and velocity into the parameter set of the basic personal space, particularly \u03c3 py i . We divide the space around the human pi into two parts: (1) frontal area, and (2) back area, as seen in Fig. 2(b). As humans mostly move forward rather than backward, we propose new factors fv and f f ront influencing the \u03c3 py i value, which is computed as follows: \u03c3 py i = { (1+ vp i fv + f f ront)\u03c3 py 0 if frontal area \u03c3 py 0 if back area (8) Note that (x,y) is the coordinate of a point around the person pi, vp i is velocity and fv is its normalization factor, f f ront is a factor proposed for the frontal area of the human, and \u03b8 p i is the human orientation. Equation (8) returns the standard deviation \u03c3 py i , which is a part of the parameter set [Ap,\u03c3 px 0 ,\u03c3 py i ], for the Gaussian function f eps i (x,y) that represents the EPS around the person pi"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure12.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure12.3-1.png",
        "caption": "Fig. 12.3 Example of the \u201ccycle\u201d issue during computation using the Newton\u2019s method",
        "texts": [
            " ; x N \u0394x J , and f i x \u00bc \u2202U x \u2202xi U x1; . . . ; xi \u00fe \u0394x f ; . . . ; xN U x1; . . . ; xN\u00f0 \u00de \u0394x f \u00f012:30\u00de where \u0394x J and \u0394xf should be very small, yet they should also be chosen to be sufficiently large and with signs (positive or negative) such that the calculated Jij would have finite and representative values. There are special cases that the Newton-Raphson method may not be able to find an optimal solution. Sometimes, the estimates may get stuck in a \u201ccycle.\u201d Considering a one-dimensional example as we can see from Fig. 12.3, the function has the same derivative in xA and xB. The Newton iterative scheme cycles between these two values xA and xB never induce convergence to the solution xS. One way to handle this \u201ccycle\u201d issue is to alter the recursive scheme. The inverse of Jacobian matrix J 1(x o) in Eq. 12.29 has a magnitude that can directly induce the solution if f(x) is linear. However, this magnitude does not apply for a nonlinear f(x). We may modify Eq. 12.31 as x n\u00fe1 \u00bc x n KdampJ x n 1 f x n , for n \u00bc 0,1,2, "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002195_tmag.2014.2317491-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002195_tmag.2014.2317491-Figure9-1.png",
        "caption": "Fig. 9. Cableless magnetic actuator system.",
        "texts": [],
        "surrounding_texts": [
            "Based on the above-mention measurement results, additional devices of a CCD camera, two LED lights, a transmitter with a wire antenna, and various types of batteries were attached to the proposed cableless magnetic actuator, as shown in Figs. 9 and 10. These additional devices are commercial product. Permanent magnets A and B used in the propulsion module are the same as those shown in Fig. 1. The CCD camera is a cube with a side length of 8.5 mm and a total mass of 1.2 g. The transmitter has a width of 9 mm, a height of 4 mm, a thickness of 0.5 mm, and the total mass of the transmitter, including the wire antenna is 0.7 g. The cableless actuator system is 95 mm in length and 12 mm in diameter and has a total mass of 18.5 g. This actuator system is capable of moving in pipes of 12.5 mm in diameter. Based on the results shown in Figs. 6 and 7, seven SR920W button batteries were used to drive the propulsion module of the actuator for the purpose of light-weight and drive of long distances. An electric current can decrease by reducing the number of button battery. The consumption current in the case of seven button batteries was 20 mA. However, that of ten button batteries was 28 mA. In addition, three Maxell SW43 button batteries were used to drive the CCD camera and the transmitter and two Maxell SR521SW button batteries were used to power the two LED lights. Both of these batteries (SW43 and SR521SW) have a nominal output of 1.55 V. These batteries were arranged in series. For the purpose of shortening the length of the actuator system, four reed switches were located outside the propulsion module, as shown in Figs. 9 and 10. Fig. 11 shows the ON\u2013OFF areas of reed switch pairs A and B that form around the two ring type permanent magnets. In this actuator system, the X positions of pairs A and B are \u22125 and 5 mm, respectively, and the Y positions are 4.5 and \u22124.5 mm, respectively. Consequently, the duty factor of the square wave produced by this inverter module was 25%. Fig. 12 shows the relationship between the tilt angle \u03b1 and the speed of the actuator system in a pipe with an inner diameter of 12.5 mm for the case in which the supporting forces of the actuator system are 0.12, 0.24, and 0.4 N. The supporting force was changed by varying the width of the rubber legs as 6.6 and 7.9 mm. As the supporting force increases, the dependence of the speed of the actuator system on the tilt angle decreases. Table I shows the continuous motion time and range of the actuator system for supporting forces of 0.12, 0.24, and 0.4 N. When the supporting force of the rubber legs was 0.24 N, the results demonstrated that the actuator system was able to inspect pipe at a vertical upward speed of 87.1 mm/s. The continuous motion time for the case in which seven batteries were used to drive the propulsion module of the actuator system was 50 min, and the CCD camera and the transmitter were operated for 52 min using three batteries. Accordingly, this actuator system can inspect the conditions of the inner walls of pipe for over 50 min. For example, the total range of this actuator system was 402 m in the horizontal direction and 261 m in the vertical direction for the case in which the supporting force of the actuator system was 0.24 N. Fig. 13 shows an image of a system that is capable of thin pipe inspection. This actuator system was confirmed to enable inspection of thin pipe using only a 2.4 GHz receiver and a PC with video-capture software. Since the maximum distance of the transmitter attached to the actuator system is 30 m, the inspection range is limited to <30 m. Nonmagnetic materials having superior corrosion properties, such as stainless, aluminum, and copper are often used in pipes in chemical plants. The proposed actuator system was confirmed in a previous study to be capable of reversible movement [8]."
        ]
    },
    {
        "image_filename": "designv11_64_0002935_gt2016-56508-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002935_gt2016-56508-Figure3-1.png",
        "caption": "Figure 3: AFB assembly",
        "texts": [],
        "surrounding_texts": [
            "This paper presents the analytical and experimental rotordynamic studies of a shaft with a large overhung mass supported by air foil bearings. The bearing coefficients are calculated by adopting the perturbation method. The performance quality of the radial air bearings are individually checked through the lift-off test. The rotordynamics of the rotor-air bearing system is evaluated analytically through the synchronous imbalance response studies, in which the system is modeled with three different approaches: non-linear rigid shaft, linear flexible shaft, and non-linear flexible shaft. The simulated imbalance responses show a stable steady-state vibration up to 200krpm. The high speed bench test results show stable steady-state vibration between 100 and 160krpm with a very small sub-synchronous vibration with frequency locked to the system natural frequency."
        ]
    },
    {
        "image_filename": "designv11_64_0001168_gt2014-26128-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001168_gt2014-26128-Figure3-1.png",
        "caption": "Figure 3: Image showing the back chamber and planes of comparison, shown in black",
        "texts": [
            "6 million cells took approximately 4 hours to converge on 16 Intel Xeon E5472 3.0GHz CPUs. Mesh Independence A mesh independence study was conducted to identify an appropriate mesh density for the calculations. Five meshes were investigated ranging from 0.32 million cells to 1.6 million. The meshes are tabulated against their identifier in Table 1. In order to compare data for the five meshes average velocity profiles were created on the centreline of places at the shroud outlet and 3 vertical planes within the back chamber. The locations of these planes are depicted in Figure 3, and profiles of velocity magnitude for the 5 meshes at Position 1 can be seen in Figure 4, for a shaft speed of 5,000 rpm. 3 Copyright \u00a9 2014 by Rolls-Royce plc Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use The mean velocity magnitude for each mesh at Position 1 is presented in Table 2 along with the percentage difference from mesh M1(the finest mesh) and an L2 measure. The L2 measure is calculated as L2 measure = \u221a \u2211(\ud835\udc62\ud835\udc401 \u2212 \ud835\udc62\ud835\udc40\ud835\udc5b) 2 \u2211\ud835\udc62\ud835\udc401 2 ",
            "org/about-asme/terms-of-use Data from the single phase model showed that there would be high shear acting on wall films within the back chamber. This was as expected as the chamber is bounded front and back by two highly rotating walls. As a result it was thought that the flow would be largely shear dominated and that gravitational effects would be secondary. This facilitated the use of the axisymmetric segment modelling approach. The velocity profile on a radial line approximately equidistant from front and back walls (plane 3 as illustrated in Figure 3) is shown on Figure 12 (blue \u201cx\u201d symbols).where the velocity gradient is seen to be quite steep at the outer radial positions. When the oil phase is added to the model however, the core velocities in the back chamber drop significantly and wall shear is no longer as highly dominant. The velocity profile from the Eulerian simulation is on Figure 12 as red \u201cplus\u201d symbols. It is clear that shear acting on wall films is much reduced in the two-phase case. Considering the increased weight of oil, as the film develops body forces become increasingly important and it no longer becomes acceptable to neglect gravity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001589_0309364613498333-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001589_0309364613498333-Figure1-1.png",
        "caption": "Figure 1. Late stance.",
        "texts": [
            " While I agree that further research is needed on the effect of tuning, I am very concerned that the conclusions listed in the article might dissuade clinicians from undertaking tuning of rigid AFOs in the meantime. When I first described the tuning process with AFOs for young children with cerebral palsy (CP) (actually, I called it \u2018fine-tuning\u2019 which perhaps more closely describes the sensitivity of the process), one of the interesting outcomes was facilitation of stability in mid to late stance.1 In normal gait, this is achieved by forward inclination of the thigh and alignment of the ground reaction force (GRF) in front of the knee joint centre and, crucially, behind the hip joint centre (Figure 1). Thus, stabilising external extension moments are applied at these joints, and the second peak of the vertical component of GRF is greater than body weight. Research has shown that reduced second peak of GRF, and thus instability, in CP is a common problem.2 In my opinion, when using AFOs, one objective is to achieve, as closely as possible, similar segment, joint centre and GRF alignments as in normal gait. From my research and clinical experience, it is clear to me that this process requires correct AFO design and fabrication \u2013 accommodating gastrocnemius contracture/tone, adequate stiffness, choice of initial shank to vertical angle (SVA) and so on \u2013 followed by tuning, in particular alteration of the SVA by the addition of small heel wedges between AFO and shoe"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002624_jae-150151-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002624_jae-150151-Figure1-1.png",
        "caption": "Fig. 1(a). The physical model of the three-phase synchronous generator.",
        "texts": [
            "\u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 i1r = i1r+ + i1r\u2212 = i1r+e j\u0394r + i1r\u2212e\u2212j\u0394r = \u221a 2 3( i1ra + i1rb + i1rc) = \u221a 2 3 [ b0r b 1 r b 2 r ] \u23a1\u23a3 i1ra+(t) i1rb+(t) i1rc+(t) \u23a4 \u23a6+ \u221a 2 3 [ b0r b 1 r b 2 r ] \u23a1\u23a3i1ra\u2212(t) i1rb\u2212(t) i1rc\u2212(t) \u23a4 \u23a6 \u0394r = \u222b t 0 \u03c9rdt (2) Here, br is the spatial operator, and br = ej120 \u25e6 . The equivalent three-phase rotor current fundamental vectors are i1ra, i1rb, i1rc respectively, the instantaneous value of the three-phase rotor positive current fundamental vectors are i1ra+(t), i1rb+(t), i1rc+(t) respectively, the instantaneous value of the threephase rotor negative current fundamental vectors are i1ra\u2212(t), i1rb\u2212(t), i1rc\u2212(t) respectively, and the rotor electrical angular speed is \u03c9r. Figure 1(a) could be deemed as a two-pole generator or a pair of magnetic poles in the multipolar electric machine. When the magnetic field effect of the rotating rotor is perceived as the magnetic field effect of the static three-phase windings, the equivalent static rotor model for the generator is displayed in Fig. 1(b). Through the Eqs (1) and (2), the physical model in Fig. 1(a) could substitute for the equivalent model in Fig. 1(b), in which the stator and rotor are both static. The synchronous generator could be viewed as two windings rotating at the synchronous speed and the another two windings rotating at the reverse synchronous speed. The interaction between the stator magnetic field and the rotor magnetic field is described by interaction of the four current vectors in Fig. 1(c), which include the stator positive current fundamental vector and the rotor positive current fundamental vector at the synchronous speed, and the stator negative current fundamental vector and rotor negative current fundamental vector at the reverse synchronous speed. Among them, i1+ is the resultant positive current fundamental vector from the stator positive current fundamental vector and the rotor positive current fundamental vector, and i1\u2212 is the resultant negative current fundamental vector from the stator negative current fundamental vector and the rotor negative current fundamental vector"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure2-1.png",
        "caption": "Fig. 2. Quarter of a 2-D model of the squirrel-cage induction motor.",
        "texts": [
            " (23) This corresponds to a system247 [Sk ]ak = bk (24) of linear algebraic equations, where the degrees of freedom are248 organized into a vector ak , and the elements of the system matrix249 [Sk ] and the input bk vector are250 Sk ji = \u222b \u03a9 \u03bc\u22121 0 \u2207\u03bbi \u00b7 \u2207\u03bbj da (25) bk j = \u2212 \u222b \u03a9k (Mk \u00d7\u2207\u03bbj )z da + \u222b \u03a9k Jk\u03bbj da. (26) The linear system matrix [Sk ] is the same for all k; hence,251 a single matrix decomposition enables the computation of all252 pairs (Hk ,Bk ).253 V. EXAMPLES254 In this section, decomposition of the magnetic field and torque255 is demonstrated with a squirrel-cage induction motor and a per-256 manent magnet synchronous motor.257 A. Squirrel-Cage Induction Motor258 Fig. 2 shows a quarter of the model geometry of the squirrel-259 cage induction motor. The essential motor parameters are in260 Table I, and the BH-curve of the iron cores and shaft are in261 Fig. 3. The iron cores are assumed nonconducting, whereas262 the shaft and squirrel-cage have otherwise cite it at appropriate263 place. a conductivity of 4.3e6 S/m and 3.2e7 S/m, respectively.264 Each stator coil is connected to a balanced three-phase voltage265 Vq = V0 cos(2\u03c0ft + q2\u03c0/3), where f = 50 Hz, q \u2208 {0, 1, 2}266 and V0 = 400 V"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003238_detc2016-59654-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003238_detc2016-59654-Figure6-1.png",
        "caption": "Figure 6 10-minute fatigue damage at recess point of sun gear",
        "texts": [
            " Therefore, the sun gear is most critical in terms of the contact fatigue for this drivetrain [14]. Using the procedure summarized in Section 4, the 10- minute fatigue damage 10min 10 10( , , )D v id at the recess point on 6 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90698/ on 04/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use the most critical tooth of the sun gear is calculated for each wind load scenario defined by V10 and I10 as shown in Fig. 6. Each bar in this figure indicates the magnitude of the 10-minute fatigue damage. It is observed from this figure that the fatigue damage increases with an increase of the mean wind speed since the number of load cycles and contact loads increase as the rotor speed increases. However, the fatigue damage plateaus when the wind speed gets higher than the rated speed of 16 m/s. This is attributed to the fact that the blade pitch control is activated at the rated speed (16 m/s) to maintain constant power generation at a constant rotor speed [40]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002377_compel-06-2015-0217-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002377_compel-06-2015-0217-Figure1-1.png",
        "caption": "Fig. 1: 3DOF Helicopter.",
        "texts": [
            " The simulation results of the suggested controllers are given in section V. Some experimental results of the proposed controllers are shown in section VI. Finally, conclusion of this work is presented in section VII. The system modeling remains an essential task before the control design phase. Indeed, the ability to describe and explain the various phenomena involved and interacting in the helicopter dynamics has a large impact in practice. A 3-DOF helicopter product by Quanser (QUANSER, 2009) is considered, as shown in Figure 1. The helicopter is a rigid body fixed at a spherical joint to a suspension point around which it can turn freely in any direction, and has two control inputs in the form of individually controllable motors leaving a degree of freedom unactuated. Motors are fixed symmetrically with respect to a side of the body. At the opposite end, a counterweight is attached to reduce the effect gravitational forces. The elevation angle, the pitch angle and the travel angle are called \u03b8, \u03d5 and \u03c8, respectively. D ow nl oa de d by U ni ve rs ity o f Sy dn ey L ib ra ry A t 0 6: 22 2 2 M ar ch 2 01 6 (P T ) 3 The elevation torque is controlled by the total force Fm = Ff + Fb"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure16-1.png",
        "caption": "Fig. 16. Sliding Shoe Analysis",
        "texts": [
            " He disagreed with the author\u2019s statement that the sliding shoe tended to have a lower output than the pivoted type and to be more stable, but that it is possible to obtain similar performance with either type. The sliding shoe scored on all counts. The at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from 51 DISCUSSION ON INTERNAL E X P A N D I N author\u2019s conclusions were reached because of an error in calculations, and because the premises on which comparisons were based, were not truly representative. The error in calculations could be, perhaps, best explained with reference to Fig. 16, which showed the particular sliding shoe which the author selected for a basis of comparison. The locus of the true centre of pressure is shown in heavy line, and the author\u2019s approximation is chain dotted. I t would be seen that, for normal operation, namely, when p = 0.35 and thereabouts the approximation was very good indeed. But for large values of friction-for example, to determine spragging, which was one of the author\u2019s comparators-the differences were serious. For the case shown, according to the author\u2019s method spragging took place when p = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003134_978-981-10-2875-5_70-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003134_978-981-10-2875-5_70-Figure3-1.png",
        "caption": "Fig. 3 Schematic diagram of the minimum composite unit",
        "texts": [
            " Then, the geometric characteristics of the Si can be expressed as: S1 k S2 k S3?AB, S4 k S5 k S6?AC, S7 k S8 k S9?BC, S10 k S11?AH, S12 k S13?CH, and S14 k S15?BH. Besides, the S11, S13 and S15 Fig. 2 Schematic diagram of the tetrahedral element are coplanar and perpendicular to the planes AOH, COH and BOH, respectively, where the point O represents the circumcenter of the triangle ABC. The minimum composite unit of the deployable truss antenna is constituted by three tetrahedral elements, as shown in Fig. 3. Then the whole deployable truss antenna shown in Fig. 1 can be obtained by extending the minimum composite units from the central axis of the antenna. The reflecting surface of the antenna spliced by the minimum composite units is shown in Fig. 4, in which the triangle planes are the undersurface of the tetrahedral elements and the color hexagons represent the minimum composite units. In order to analyze the DOF of the tetrahedral element, it is firstly divided into two parts in this section, as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000212_978-981-15-5712-5-Figure24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000212_978-981-15-5712-5-Figure24-1.png",
        "caption": "Fig. 24 The sudden change in the growth rate of iron production",
        "texts": [
            " In certain areas of mechanical engineering, chemistry started playing a major role and this was the impetus for the development of chemical engineering. Aircrafts and space devices coming into the picture in a major way caused a further split of mechanical engineering, and the result was the emergence of aerospace engineering. It has been mentioned earlier that the invention of prime movers led to an Industrial Revolution\u2014called the First Industrial Revolution. If production of iron is taken as a measure of the degree of industrialization, then Fig. 24 shows the characteristic change in the growth rate of Iron production. The sudden change in the slope indicates a paradigm change and the Industrial Revolution. The Second Industrial Revolution (IR) is considered to be caused by the emergence of computer technology and the semiconductor industry. This can be identified with the help of the economic growth characteristics shown in Fig. 25. The primary impetus to the Second Industrial Revolution came from the miniaturization of electronic devices",
            " But first, we should take note of the fact that both series and parallel versions have the same mechanical construction and that there is a short flexure that enables the wide beam to rotate. Also to be noticed is how parallel electrical connection could be given to a heatuator in the layout of microfabricated components. As in the bent-beam actuator, many series heatuators can be made into an array to generate large force [12]. But we can do more with the parallel heatuators. This is shown in Fig. 24. In Fig. 24a, we see an expanding actuator that has four parallel heatuators. When voltage is applied between its two anchor pads at the bottom, they remain stationary but the movable pad at the top moves upward. Notice how the wide and narrow beams are arranged in order to achieve the upward or expanding motion, which can be seen in Fig. 24b. On the other hand, the narrow and wide beams are switched, and the heatuators on either side are kept at an angle, as shown in Fig. 24c. Now, we see a contracting actuator as can be discerned from Fig. 24d. The possibilities for design are almost unlimited with electrothermal actuation. It paves the way for topology optimization involving multiphysics simulation [31\u2013 33]. Analysis of electrothermal actuators is challenging because three simulations\u2014 electrical, thermal, and elastic\u2014are to be performed in a sequence. Conduction, 48 G. K. Ananthasuresh convection, and radiation effects as well as temperature-dependent properties are to be accounted for in simulation and design [34]. Both intuition and systematic optimizations can lead to interesting designs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001400_mecbme.2014.6783284-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001400_mecbme.2014.6783284-Figure1-1.png",
        "caption": "Figure 1 A Schematic of the Biped Robot",
        "texts": [],
        "surrounding_texts": [
            "particular has stirred a great deal of interest in the last few years. In this paper, we focus on the energetic aspect of a bipedal robot with a torso walking on level ground.\nI. INTRODUCTION\nA popular theory regarding legged locomotion is that human (or animals) move in a way that minimizes energy in one form or the other [1-4]. Inverted pendulum walking is one of the idealized gaits that can be used to describe walking. In an inverted pendulum walking the hip moves in series of circular arcs, vaulting over a straight leg. The transition from one circular arc to the next is realized via impulsive push-off [4] that happens just before heel strike. There seems to be very little work in the literature that investigated the role of the upper body in walking and running[5]. This ongoing work looks at the \"optimal\" postural strategies and energetics of a simple biped with a torso to adapt to uphill and downhill walking as compared to level walking for a given forward velocity in terms of say stride length and tilt of upper body. The results reported in this particular paper are only for the case of level ground.\nII. MATHEMATICAL MODEL AND PROBLEM FORMULATION\nWe treat the body as a hip mass mH at a position (xH,yH) at time t, and a trunk mass mT at a position (xT,yT ). The trunk or torso is controlled via a torque between the stance leg and the torso. The legs are massless. The fluctuations of the leg length q(t) due to flexion of the hip, knee and ankle are incorporated in a single telescopic axial actuator that carries a compressive time varying force F(t). The leg has a maximum allowable leg extension, such that\n)(22 tqRyx HH , where R is the nominal length of the\nleg. It is assumed that during the stance phase, the foot in contact with the ground does not slip, and at most one foot can be in contact with the ground at any given time, and that there is no flight phase. The left and right legs have identical force and length profiles.\nA gait is characterized by the position and velocity of the hip mass and torso mass, by the step period and by F(t) and the torque (t) and the maximum allowed leg extension. Using Lagrange formulation the equations of motion are:\n*Research supported by Sultan Qaboos University under code number IG/ENG/MIED/12/01. Dr. A. S. AlYahmedi is with the Department of Mechanical & Industrial Engineering, College of Engineering, Sultan Qaboos University, P.O. Box 33, Al-Khod, Muscat, 12, Sultanate of Oman (phone: 968-95209194; fax: +968-2414-1316; e-mail: amery@squ.edu.om).\nM. A. Sayari was with National School of engineering of Sfax.\n(e-mail: sayariaminemm@gmail.com).\n(1.a)\n(1.b)\nWhere, L is distance from hip to the center of mass of the torso, and g is the gravitational constant. To reduce the number of parameters scaling was used. Scaling shows its effectiveness in gait analysis [6] and meant that the equations are not anymore belonging to any dimension thus the person could be treated as an ensemble of ratios.\nAfter scaling the equation of the model using\n, and . The only two free parameters\nremaining are and . The optimizer will seek solutions as\ntwo parameters are varied, namely, normalized speed , and normalized step length . The two remaining parameters,\nnamely will be fixed. The governing equations of\nthe motion after non dimensiolizing are\n(2)\nWhere,\nare the non-dimensional variables.\n978-1-4799-4799-7/14/$31.00 \u00a92014 IEEE 382",
            "Figure 3 Fluctuation of the non-dimensional leg length\nWhere,\nThe humans have an ideal energetic walk. This idealized posture is the result of a natural optimization of the motion of the different parts of the body. This shows a close relation between optimization and walking. To highlight this close relation, numerical optimization has been applied to the mathematical model of the human. The optimal solution have cost arbitrary close to zero unless the optimization is constrained. We set the maximum length to\nrepresenting a maximum increase of 10% of the\nnominal leg length. The optimization is done for a given step length, d. Further there is no cost associated with standing so we optimize for a given average speed v.\nGiven a particular d, assuming that a given step starts with the nominal leg length (R+q(0) = R); we seek the control strategy ; that minimizes the cost of transport [7]\n(3)\nSubject to Eq. (2) and which satisfy the constrains of\nperiodicity: same position and velocity of the torso before\nand after the stride , Same velocity of\nSame position of the hip in Y direction before and after the stride\n(4.3)\nAnd stride length, D\n(4.4)\nAnd the one on the maximum extension of the leg\n(4.5)\nWith (4.6)\nAlthough a wide range of gait patterns are possible; people usually prefer just two; walking at low speed and running (bouncing up off leg between flight phases at higher speeds,\n[7]). A simulation for an average person [9] with\nhave been done.\nFor each and within the range of walking [10], multiple optimizations runs, each one started with different initial conditions and converges to a unique solution which determines the optimal gait for this speed and step length. Indeed running the model with piecewise constant controller\nfor and gives the results shown in Fig.2-Fig.5\nAfter heel strike the upper body tends to rotate slightly\nbackward over the stance leg to absorb the impact followed\nFigure 2 Torso angle"
        ]
    },
    {
        "image_filename": "designv11_64_0000464_b978-0-08-096532-1.00508-2-Figure17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000464_b978-0-08-096532-1.00508-2-Figure17-1.png",
        "caption": "Figure 17 Various designs of grain selector for single-crystal turbine blades: (a) angled; (b) spiral (helix); and (c) restrictor. (After Goulette, M. J.; Spilling, P. D.; Arthey, R. P. Cost Effective Single Crystals. In Superalloys; AIME: Warrendale, PA, 1984.). Reproduced from Dai, H. J.; D\u2019Souza, N.; Dong, H. B. Grain Selection in Spiral Selectors during Investment Casting of Single-crystal Turbine Blades: Part I. Experimental Investigation. Metall. Mater. Trans. A 2011, 42A (11), 3430\u20133438.",
        "texts": [
            " Improvement of creep ductility and thermal fatigue resistance by orientating the preferred <001> crystallographic growth direction, which coincides with the elastically soft direction, parallel to the maximum loading direction (64\u201366). In order to grow single-crystal component, a grain selector is implemented into the directional solidification process in a modified Bridgman furnace to ensure that only one grain can survive in the final structure (Figure 16) (25,26). Different single crystal (SX) selector designs have been employed (Figure 17) (26,67) to select a single grain. The restrictor selector was found to be less efficient since a longer length is required to select a single grain. In the angled selector, problems occur around the corners because new grains often nucleate due to the sudden change in growth direction and the high temperature gradient (33,68). The most commonly used grain selector is the spiral selector. The spiral grain selector consists of two parts (Figure 18): a starter block referring to competitive growth for the grain orientation optimization and a spiral grain selector facilitating dendrite branching to ensure that only single grain eventually survives at the top of the seed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003201_icpds.2016.7756677-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003201_icpds.2016.7756677-Figure1-1.png",
        "caption": "Fig. 1. Construction of the TEFC induction motor (1 \u2212 shaft, 2 \u2212 rotor winding end-ring, 3 \u2212 frame, 4 \u2212stator iron, 5 \u2212 rotor iron, 6 \u2212 end part of stator winding, 7 \u2212 fan)",
        "texts": [
            " THE FORMATION OF GENERALIZED THERMAL MODELS To build a simple model for thermodynamic modeling requiring minimal quantity of information it is necessary for the beginning to define which factors, determining the model adequacy, are most substantial. To do this it is necessary to find out details of heat processes proceeding in big number of different motors. It demands a great effort and time expenses. Also in such case it is difficult to discover substantial regularities in a large volume of data. On this reason on the first step of the work was done an attempt to make a clustering of induction motors belonging to the widely used in Russia type 4A with mesh winding. The construction of such TEFC motor is shown in Fig. 1. Clustering was carried out for set of 82 motors A4 with power from 60 W to 90 kW. The set was divided on two groups of motors with stator winding insulation class B (50 motors) and class F (32 motors). As input parameters was used the vector containing 19 components. This vector includes parameters of 5-node thermodynamic model of the TEFC motor on a per-unit basis (losses, heat capacities and thermal conductivities). As a measure of clusters similarity the sum of absolute differences was used (this approach is called \u201c\u0421ityblock\u201d or \u201cManhattan distance\u201d) [12, 13]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.7-1.png",
        "caption": "FIGURE 8.7",
        "texts": [
            " These equations are written for the general motion of individual bodies with the addition of constraint conditions. Usually the equations of motion are derived from Newton and Euler equations or from Lagrange equations. We mentioned Newton\u2019s laws and Lagrange\u2019s equations; now we briefly discuss Euler equations, which govern the rotational motion of rigid bodies. A rigid body rotates together with its body-fixed reference frame x-y-z at an angular velocityuwith respect to the inertia (fixed) frame X-Y-Z, as shown in Figure 8.7. Note that we assume that the bodyfixed reference frame also originates at the mass center of the rigid body. For a rotating body, the angular momentum Ho is most conveniently expressed in the body reference frame x-y-z, where point o is the origin. Ho can be written as Ho \u00bc Hxi\u00fe Hy j\u00fe Hzk (8.55) where the unit vectors i, j, and k align and rotate with the x-, y-, and z-axes, respectively. From Eq. 8.29, we have Mo \u00bc _Ho (8.56) where Mo is the vector of external moment applied to the body referring to the body reference frame x-y-z"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.2-1.png",
        "caption": "Fig. 10.2 Design for forging on the resultant aspect ratio",
        "texts": [
            " Such shrinkage can be eliminated by reducing the number of walls and avoiding sharp convex/concave angles in the cast. That is, chamfer or fillet edges are preferred. For fillet edges, it is conservative to maintain the fillet radius to be between one-half and one-third of the section thickness to prevent the formation of shrinkage cavities. This criterion also facilitates smooth flow streamlines of molten metals during processing. In design for forging, the fundamental principle is minimizing required deformation during processing. An example of this design consideration is shown in Fig. 10.2. Let\u2019s imagine that a rectangular block of material is forged to have a cross-shaped cross section with one direction longer than the perpendicular direction in the cross section. It is better to have the cavities on the forging dies to be shallower, rather than deeper. It is because we may consider the aspect ratio of cavity shape should roughly represent the required material strain for permanent deformation. The local spatial strain level should preferably be in the perfect plasticity region or the strain hardening region"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001399_1464419315571983-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001399_1464419315571983-Figure4-1.png",
        "caption": "Figure 4. Changes of the angular position and contact load of the rolling element due to the orbital motion.",
        "texts": [
            " This misalignment results in the unbalance of component in the system, and then centrifugal force will emerge when the rotor rotates with a certain speed. The centrifugal force of each rotational component is a harmonic excitation force to the rotor system and can be written as Fcx\u00f0t\u00de \u00bc me! 2 s e i\u00f0!s t\u00fe 0\u00de Fcy\u00f0t\u00de \u00bc ime! 2 s e i\u00f0!s t\u00fe 0\u00de ( \u00f013\u00de Various critical speeds of the rotor-bearing system can be obtained by calculating the unbalance response of the system under different rotational speed and the response\u2013speed curve can be plotted. As shown in Figure 4, nonuniform load distribution among the rolling elements will occur when a radial static load acting on the bearing. If the static bearing displacement is supposed to be invariable when the rolling elements pass the bearing chamber with a constant orbital speed, the angular position and contact load of each rolling element will vary periodically and the period fluctuation of the bearing stiffness will be caused. This parametric excitation is called VC effect and it can also result in a vibration of the system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002357_978-3-319-17067-1_24-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002357_978-3-319-17067-1_24-Figure2-1.png",
        "caption": "Fig. 2 The leg mechanism construction: a sketch diagram, b 3D module, c prototype",
        "texts": [
            " Meanwhile, the foot trajectory, generally divided into the \u201csupport phase\u201d and the \u201ctransfer phase\u201d, should meet the functional requirements, which can be summarized as follows. The desirable foot trajectory is shown in the Fig. 1: 1. The trajectory must be a closed symmetrical curve without intersections. 2. The support phase of the foot-point is approximately a straight line. 3. The height of the stride H in the transfer phase is as high as possible, and the length L in the support phase is as long as possible to ensure the walking ability. Based on the functional requirements, the 1-DoF planar leg mechanism with the close-chain full-pivot feature is presented in Fig. 2. The linkages in Watt chain [15] are designed as ground link, crank, thigh link, and shank link, respectively. In Fig. 2a, the parameters are described as: ri (i = 1, 2, \u2026, 9), the length of linages; \u03b8i,j (i = 1, 2,\u2026,9; j = 1, 2, \u2026,9), the angle between the linkages i and j. The design parameters are listed in Table 1. The arrangement of the DQV is shown in Fig. 3a. The whole mechanism consists of two identical quadruped mechanisms on each side of the frame. In each quadruped mechanism, the cranks arranged in the front and rear legs on the same side have a zero radian phase difference, and the cranks arranged in the left and right legs have a 180\u00b0 phase difference"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003343_cefc.2016.7816296-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003343_cefc.2016.7816296-Figure1-1.png",
        "caption": "Fig. 1.Topology of the OR-FSPM machine.",
        "texts": [
            " At the same time, the FSPM machine with FSCWs are popularly used in EVs, for their simple topology, strong robustness, shorter end-winding length, strong flux weakening ability[1, -2]. In this paper, a 3-phase, 48/56 stator/rotor poles OR-FSPM machine with FSCWs is proposed and modelled by FEA. Comprehensive simulations indicate that this structure has less total harmonic distortion (THD) of back-EMF, higher average torque, lower torque ripple and high efficiency, which is a good candidate for in-wheel motors based EVs. The topology of the proposed OR-FSPM with FSCWs adopted in the stator is shown in Fig. 1. It can be seen that its structure is double salient, and has passive rotors without windings or PMs, which are identical to those of switched reluctance machines. Furthermore, this can produce sinusoidal back-EMF curve with little harmonic components due to its operation principle and FSCWs. This work has been supported by National Natural Science Foundation of China (NSFC), 51377065 and 61301035, and Hubei Province Science and Technology Supporting Program 2014BAA035. The curves of back-EMF, A-phase flux linkage and output torque are shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.35-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.35-1.png",
        "caption": "FIGURE 6.35",
        "texts": [
            " The steering rack part is connected to the vehicle body by a translational joint and connected to the tie rod by a universal joint. The translation of the rack is related to the rotation of the steering column by some kind of coupling statement that defines the ratio; such constructs are common to most general purpose software packages. Attempts to incorporate the steering system into the simple models using lumped masses, swing arms and roll stiffness will be met with a problem when connecting the steering rack to the actual suspension part. This is best explained by considering the situation shown in Figure 6.35. The geometry of the tie rod, essentially the locations of the two ends, is designed with the suspension linkage layout and will work if implemented in an \u2018as is\u2019 model of the vehicle including all the suspension linkages. Physically connecting the tie rod to the simple suspensions does not work. During an initial static analysis of the full vehicle, to settle at kerb height, the rack moves down with the vehicle body relative to the suspension system. This has a pulling effect, or pushing according to the rack position, on the tie rod that causes the front wheels to steer during the initial static analysis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000924_amr.1059.1-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000924_amr.1059.1-Figure2-1.png",
        "caption": "Fig. 2 Scheme of equipment for testing abrasive wear on emery cloth (1 \u2013 rotating horizontal plate; 2 \u2013 emery cloth; 3 \u2013 test sample; 4 \u2013 clamping head; 5 \u2013 weight; 6 \u2013 shaft for radial shift; 7 \u2013 end switch)",
        "texts": [
            " Micro-hardness was measured using the device FM-100 (Future-Tech Corp., Japan) according to the Vickers method under load 0.981 N (100 g). Measurement was performed in cut, in the direction from the surface of hard deposition towards the base material, with spacing 0.5 mm, whereby the first measurement was in depth 0.5 mm from the surface of the deposited layer. The abrasive wear resistance of the materials tested on the emery cloth was determined on the device of pin-on-disc type, the principle of which is shown in Fig. 2. The diameter of the emery cloth was 480 mm. The sample moved along the spiral so that the whole friction path was 50 m. Sample speed was 0.5 m.s-1 and specific pressure 0.32 MPa. The emery twill A99 (Globus 120) was used as the emery cloth. Values of relative abrasive wear resistance of tested materials were calculated according to the following relation: E E abr m m \u03c1 \u03c1 \u22c5 \u2206 \u2206 =\u03a8 , (1) where: \u03a8abr relative abrasive wear resistance on emery cloth [\u2013]; Em\u2206 average weight loss of etalon samples [g]; m\u2206 average weight loss of tested material samples [g]; \u03c1E specific weight of etalon material [g"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000395_978-94-007-4132-4_8-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000395_978-94-007-4132-4_8-Figure5-1.png",
        "caption": "Fig. 5. Systematics of six-link dwell mechanisms with simple turning pairs by Kurt Hain",
        "texts": [
            " in so-called \u201cdead positions\u201d, the velocity of the driven link is also zero, but not its acceleration. The dwells generated by linkages are normally only approximate ones, i.e. the zero values of velocity and acceleration of the driven link mentioned above are reached only approximately or on an average [5, 6]. In 1980 Kurt Hain (1908-1995) edited a booklet on dwell mechanisms [7]. In the beginning of this booklet Hain dealt with the systematics of dwell mechanisms based on six links of the type \u201cWatt\u201d and \u201cStephenson\u201d and found nine structures, Fig. 5. The driving link is marked by a circle segment with one arrow, the driven link with the possibility to have a dwell position by a circle segment with two arrows. Turning pairs with a full rotation of 360 deg are marked by a double circle, those only swinging are marked by a simple circle. So, when we bear in mind Hain\u00b4s systematic overview and look at Hoecken\u00b4s six-link dwell mechanism (Fig. 6), we find out at once that Hoecken\u00b4s choice corresponds with the number 1 of Hain\u00b4s catalogue: there is a four-link crank-rocker with a two-bar E-F coupled to it"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.26-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.26-1.png",
        "caption": "FIGURE 6.26",
        "texts": [
            " In Chapter 5 the force and moment generating characteristics of the tyre were discussed and it was shown how the braking force generated at the tyre contact patch depends on the value of the slip ratio which varies from zero for a free rolling wheel to unity for a braked and fully locked wheel. In this section we are not so much concerned with the tyre, given that we would be using a tyre model interfaced with our full vehicle model to represent his behaviour. Rather we now address the modelling of the mechanisms used to apply a braking torque acting about the spin axis of the road wheel that produces the change in slip ratio and subsequent braking force. Clearly as the vehicle brakes, as shown in Figure 6.26, there is weight transfer from the rear to the front of the vehicle. Given what we know about the tyre behaviour the change in the vertical loads acting through the tyres will influence the braking forces generated. As such the braking model may need to account for real effects such as proportioning the braking pressures to the front and rear wheels or the implementation of ABS. Before any consideration of this we need to address the mechanism to model a braking torque acting on a single road wheel"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003111_icemi.2015.7494242-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003111_icemi.2015.7494242-Figure1-1.png",
        "caption": "Fig. 1. Structure of double-fed wind turbine.",
        "texts": [
            " The paper is finally completed by summarizing the work into a few concluding remarks in Section 8. There are two kind of wind turbine: horizontal axis and vertical axis wind turbines. Megawatt wind turbines are the later which have rotor, shaft and el ectri cal generator. For the double-fed wind turbine, one of most popular commercial wind turbine, gearbox is one of the cri tic al components. The function of WT gearbox is to step up the speed of rotor rotation to a value suitable for WT generator as showed in Fig 1. Gearbox designs for wind turbine applications are differ from those used in conventional mechanical machines. Normally WT gearbox is combined with three separate stages with ratios of between 1:3 and 1:5 each. Therefore overall gear ratio comes to between about 1 :31 and 1 :88. Parallel axis gears are arranged in parallel shaft or planet gear. Wind turbines with more than 450 kW rated power have integrated gearboxes with a planet gear and two normal stages or two planet gears and one normal stage"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003090_icelmach.2016.7732872-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003090_icelmach.2016.7732872-Figure2-1.png",
        "caption": "Fig. 2. Current superimposition variable flux reluctance machine with distributed winding.",
        "texts": [
            " Niguchi and Y. Ohno are with the Graduate School of Engineering, Osaka University, Suita, Osaka, Japan (e-mail: akira.kohara@ams.eng.oska-u.ac.jp, k-hirata@ams.eng.osaka-u.ac.jp, noboru.niguchi@ams.eng.osaka-u.ac.jp, yuki.ohno@ams.eng.osaka-u.ac.jp). operational principle of the CSVFRMDW are described. In addition, the N-T characteristics are computed under vector control. Finally, the effectiveness of the CSVFRMDW is verified by comparing with the CSVFRMCW in terms of torque characteristics. Fig. 2 shows the structure and winding pattern of the CSVFRMDW, which consists of a 10-pole rotor and a 48- slot stator that has a single set of windings. The winding pattern is the same as a conventional 8-pole-48-slot permanent magnet machine. A 6-phase inverter is used to operate the CSVFRMDW using 6 phases (A, B, C, D, E, and F), which corresponds to Study on a Current Superimposition Variable Flux Reluctance Machine with Distributed Winding Akira Kohara, Katsuhiro Hirata, Noboru Niguchi, Yuki Ohno T 2 sets of 3 phases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure1-1.png",
        "caption": "Fig 1 Cross section of external rotor machine with eccentric magnetic poles",
        "texts": [
            " 51307029 and No.51077023) eccentricity on the rotor losses in permanent magnet machine is obviously missing from the literatures [6, 7]. In this paper, both an analytical method and finite element (FE) model are used for calculating the additional eddy current losses caused by air-gap eccentricity inside the rotor of high speed permanent magnet machines which is rotating at 20,000rpm with 100 kW rated power. The cross section of an external rotor machine with eccentric magnetic poles is shown in Fig 1. First, the analytical calculation of additional air-gap magnetic field and eddy current loss incurred in the rotor magnet under no-load conditions is derived, which is caused by the air-gap eccentricity. The relationship between additional eddy loss and eccentricity ratio is presented. Then the finite element simulations are used to analyze the rotor loss and the eddy current distribution respectively due to different types of eccentricity (static and dynamic eccentricity) and eccentric angle. II"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure20-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure20-1.png",
        "caption": "Fig. 20. Scheme for Hub Lubrication",
        "texts": [
            " There was room within the scantlings of existhg design to incorporate an annular oil reservoir of about 10 inches diameter and 2 inches width in the case of front hubs and of greater width in the rear hubs. If such a container were filled to a level corresponding to a line from 8 to 4 o\u2019clock by means of a plug situated on that line, then ample oil would be available to lubricate the bearings. When the hub was stationary, however, the head of oil to be retained by the seal would be a minimum. The scheme was illustrated in Fig. 20. Lt.-Col. H. E. MILBURN, M.I.Mech.E., pleaded that holes should not be made in brake drums for cooling if the vehicles were to operate in countries where there were bad roads, He had had experience of East Africa, where the soil was extremely abrasive; as good roads were non-existent, vehicles were often running quite deeply in that loose soil. In wet weather it was churned into a fine grinding compound, and brakes lasted in such circumstances for about a month. One well-known vehicle sent out there in some quantities had to have a major alteration carried out so that, as nearly as possible, the brake drums were hermetically sealed to keep out that grinding composition",
            " r at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from 64 COMMUNICATIONS ON INTERNAL EXPANDING SHOE BRAKES FOR ROAD VEHICLES Type of brake If Factor F Sensitivity Lining wear = 0.35 FO4/FO.3 - 1 x = o , - r, r Two-trailing-shoe . Fixedcam . . Two-leading-shoe . Duo-servo . . Floating expander . To find the Line of action of the resultant force on the shoe, OB, the line of closest approach, was produced to E, where OE = r,, and the resultant R was then drawn to make an angle +, the angle of friction, with OE, as shown in Fig. 20. Dr. J. R. BRISTOW, A.M.I.Mech.E., wrote, with reference to brake squed, that so-called \u201cstick-slip\u201d motion was not, in general, the result of static friction being greater than kinetic friction, but was usually due to frictional force decreasing as velocity increased. The motion was more correctly described as a relaxation oscillation, and also no period of \u201crelative rest\u201d necessarily occurred during each cycle. Moreover, the \u201cstick-slip\u201d did not excite oscillations at the natural frequency of an elastic system, but quasi-sinusoidal oscillations at the natural frequency could be maintained by virtue of the fact that energy could be fed into the oscillating system if frictional force decreased as velocity increased (relaxation oscillations being absent)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000709_978-3-319-22876-1_10-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000709_978-3-319-22876-1_10-Figure6-1.png",
        "caption": "Fig. 6. The parallel robot: (a) three-dimensional model; (b) kinematic scheme",
        "texts": [
            " The motor C and corresponding normalized reduction ratio is the optimal choice. If the robot\u2019s is given and the maximum value of it does not need to be identified in some applications, the feasible range of the reduction ratio can be found through this graph. For example, the is 4m/s in an industrial application. Then the feasible range of the normalized reduction ratio could be presented as (11) And the and are presented as in Fig. 5. The actual reduction ratio can be calcu- lated based on Eq. (2). In Ref. [12], a high-speed parallel robot was proposed as shown in Fig. 6 and the inverse kinematics of this robot has been derived. On this basis, the inverse dynamics is analyzed by using the Newton-Euler approach. Due to the space limitation, the details of the equations will not be provided. Given an arbitrary position and orientation of the mobile platform, the inputs can be identified through the vector loop method. Then the angular velocity , active joints\u2019 angular acceleration , passive limbs\u2019 angular velocity , angular acceleration and center mass velocity can be sequen- tially derived"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.24-1.png",
        "caption": "Fig. 35.24 Bearing seals - lubricated.",
        "texts": [
            " Overfilling causes rapid temperature rise particularly if speeds are high. Manufacturers supply details regarding suitable weights of grease for particular bearings. Bearings can be supplied which are sealed after prepacking with the correct type and quantity of grease. Where relubrication is more frequent, provision must be made by fitting grease nipples to the housing. Grease will then be applied by a grease gun and a lubrication duct should feed the grease adjacent to the outer ring raceway or between the rolling elements. Examples are shown in Fig. 35.24. Bearings and applied technology 543 Oil lubrication is generally used where high speeds or operating temperatures prohibit the use of grease, when it is necessary to transfer frictional heat or other applied heat away from the bearing, or when the adjacent machine parts, for example gears, are oil lubricated. Oil bath lubrication is only suitable for slow speeds. The oil is picked up by rotating bearing elements and after circulating through the bearing drains back to the oil bath. When the bearing is stationary, the oil should be at a level slightly below the center of the lowest ball or roller"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000433_j.proeng.2015.07.170-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000433_j.proeng.2015.07.170-Figure5-1.png",
        "caption": "Figure 5: Model of hand and glove including flesh, bone and viscoelastic foam",
        "texts": [
            "820 m from the butt end of the stick, respectively. The tracked glove motion from the field study was fit to third order polynomials and applied to the model in 0.1 ms increments. Contact between the blade and puck, blade and ice, ice and puck, and hands and stick were managed using the surface to surface contact control in LS Dyna. The ice was frictionless, while the coefficient of friction between the blade and the puck was 0.7 [13]. The glove was modelled with three concentric regions as shown in Figure 5. The glove exterior was given an intermediate stiffness so the prescribed motion was followed while accommodating small discontinuities in the motion. The intermediate section was given high stiffness to impart structure to the glove as occurs from a bony hand. The interior region was given a low fleshy stiffness to allow the stick to move relative to the glove as occurs in play (THUMS, Toyota Motor Corporation). A summary of the properties used for the glove materials is given in Table 3. In a slap shot, the blade contacts the ice prior to puck contact"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.10-1.png",
        "caption": "Fig. 10.10 Configuration and key parameters of milling",
        "texts": [
            ", the overlapping length of the tool and the workpiece part). In the slot cutting operation, as the tool is cutting a middle part of the workpiece, the overlapping length or t is basically the tool width. The MRR for either process can then be calculated as MRR \u00bc \u03c0 D2 D02 tf 4 D D0\u00f0 \u00de ktLfeed D\u00fe D0\u00f0 \u00dev 4D \u00f010:5\u00de The multiple (n) teeth on the milling head with a diameter of D rotating with a speed of N in rpm perform multiple cuts over the material surface with a thickness of B during the milling operation. As illustrated in Fig. 10.10 (left), the direction of the rotational axis is perpendicular to the feed direction, and these two axes are parallel to the material surface. The material surface area being removed is defined by the length of cut L and the width of cut (i.e., the tooth length) as indicated in Fig. 10.10 (right). The cutting speed v is then determined by Eq. 10.1. Meanwhile, the material sample mounted on a movable stage moves toward the milling head with a feed rate f (millimeter per minute), and the feed distance per single tooth cut (Lfeed_tooth) is f/N/ n (or the feed distance per single revolution Lfeed \u00bc f/N). 278 10 Design for Manufacturing The cutting time Tcut is calculated by Eq. 10.2, whose length of initial offset position (LA) can be further estimated by this simple proof provided below1: LA \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi t D t\u00f0 \u00de p \u00f010:6\u00de where t is the width/depth of cut",
            " The angle of a tooth cutting into the material surface can be approximated by the depth of cut z and the feed distance per tooth Lfeed_tooth as tan 1(2z/Lfeed_tooth), which is also similar to the orientation of the tooth tip with respect to the milling head center with a diameter of D, i.e., tan 1(Lfeed_tooth/D), and hence, z Lfeed_tooth 2/(2D). A quick estimation of the surface roughness (\u03b4S), which is the variation of the surface height, can simply be the average z over every single-tooth feed distance, i.e., \u03b4S Lfeed_tooth 2/(4D). In fact, a more accurate estimate of the surface roughness should also consider the tooth feed and sample feed directions. Figure 10.10 shows the up-milling case in 282 10 Design for Manufacturing which both the directions are the same. On the other hand, if the milling head is rotating counterclockwise, in this case such that its direction is opposite to the sample feed direction, this is called down-milling. To consider the roughness, it is then expressed as \u03b4S L2feed tooth 4 D Lfeed toothn= 2\u03c0\u00f0 \u00de\u00bd \u00f010:11\u00de where n is the number of teeth over the milling head. For the \u201c \u201d sign mentioned here, the up-milling adopts the \u201c+\u201d sign, whereas the down-milling adopts the \u201c\u2013 \u201d sign"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.19-1.png",
        "caption": "FIGURE 6.19",
        "texts": [
            " Vehicles with active components in the anti-roll bar system might include actuators in place of the drop links or a coupling device Modelling the anti-roll bars using interconnected finite element beams. REV, revolute joint. connecting the two halves of the system providing variable torsional stiffness at the connection. Space does not permit a description of the modelling of such systems here, but with ever more students becoming involved in motorsport this section will conclude with a description of the type of anti-roll bar model that might be included in a typical student race vehicle. A graphic for the system is shown in Figure 6.19. Graphic of anti-roll bar in typical student race vehicle. Provided courtesy of MSC Software. Modelling of anti-roll bar mechanism in student race car. The modelling of this system is illustrated in the schematic in Figure 6.20 where it can be seen that the anti-roll bar is installed vertically and is connected to the chassis by a revolute joint. The revolute joint allows the anti-roll bar to rock back and forward as the bell cranks rotate during parallel wheel travel but prevents rotation during opposite wheel travel when the body rolls"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001186_icumt.2014.7002105-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001186_icumt.2014.7002105-Figure1-1.png",
        "caption": "Fig. 1. Distance and bearing of a flight leg.",
        "texts": [],
        "surrounding_texts": [
            "Mostly used functions for lateral guidance great earth circle distance, course angle and intersection point calculations. The first basic function calculates great earth circle distance and course angle between two pairs of coordinates to a very high degree of precision. The shortest distance between two pairs of coordinates is the great earth circle line that stands for the flattened shape of the earth. But when you fly from Istanbul to Ankara on the straight line connecting them, you come across with a substantial tunnel first. The shortest great earth circle route between them lies vertically above the straight route between two pairs of coordinates. This great earth circle line can be formed by separating the earth into small parts with an imaginary plane. The great circle is the intersection with the sphere of that plane containing the center of the sphere. The shortest line between two pairs of coordinates on the surface of the earth is calculated by the arc of the great circle passing through the two waypoints. Great earth circle route is recognized on a globe based on the lines of latitude and longitude coordinates. Each line of longitude or meridian represents half of a great circle. When the prime meridian and the International Date Line are combined, they create a full great circle route which cuts the Earth into equal halves. The only line of latitude, or parallel, identified as a great circle is the equator since it passes through the center of the Earth and divides the Earth into two halves. Lines of parallels except the equator are not great circles, because when we approach near the poles, their length decrease. Great circle routes have been an important part of navigation and knowledge of them is very important for long distance travel across the world. Since great circle lines are the most efficient way to move across the globe, they are used especially in long distance travels. The datum is the mathematical model for the shape of the earth used in drafting a map. You can see the effect of datum changes by computing a conversion, then clicking different datum selections. Grid coordinates can change by several hundred meters. Since the earth is not an exact sphere, there are small errors when spherical geometry is used; the earth is approximately ellipsoidal with a radius about 6,378km according to the WGS-84 datum model. [6] 978-1-4799-5291-5/14/$31.00 \u00a92014 IEEE 214 The inputs of the main lateral guidance function are the position coordinate values of the initial and destination waypoints of the flight leg. The longitude and latitude components specify the position of any location on the planet. A flight leg is a route between two combined waypoints. The outputs of the main lateral guidance function are the great earth circle distance of the flight route and the course angle of the flight path. [7] First the function finds the unit position vectors of two pairs of coordinates in Earth-Centered Earth-Fixed coordinate system. Then the cross multiplication of two vectors is calculated. Finally, the course angle and the great circle distance between two pairs of coordinates can be calculated. P1 = cos(L1) cos(\u03bb1)i+ cos(L1) sin(\u03bb1)j + sin(L1)k (1) P2 = cos(L2) cos(\u03bb2)i+ cos(L2) sin(\u03bb2)j + sin(L2)k (2) Here, L1 and \u03bb1 are the longitude and latitude coordinates of the first waypoint. L2 and \u03bb2 are the longitude and latitude coordinates of the target waypoint. By using the two equations above the great earth circle distance between the two combined waypoints is calculated as shown below. . dist12 = tan\u22121 ( |P1 \u00d7 P2| P1 \u00b7 P2 ) (3) The bearing angle between the two pairs of coordinates is calculated by using the formula below. \u03b7P1P2 = P1 \u00d7 P2 |P1 \u00d7 P2| \u03b7P2P1 = P2 \u00d7 P1 |P2 \u00d7 P1| \u03c812 (4) = tan\u22121 ( \u2212\u03b7P1P2 P1\u03b7P1P2 \u2212 P1\u03b7P1P2 ) (5) The great circle distance and course angle of the flight route are shown in Fig. 3. The second common function used for the lateral navigation algorithm is the position calculation function. This function finds the longitude and latitude position coordinate values of the target waypoint from a given position vector, great circle distance and course angle values. For calculating the target waypoints longitude and latitude coordinates, primarily tangent unit vector in the course angle direction to great earth circle route is found. L1 is the latitude coordinate value of the first waypoint, \u03bb1 is the longitude coordinate value of the first waypoint. \u03a812 is the input course angle. The position coordinates of the target waypoint is found by the rotation of the initial waypoint and the unit vector. U\u03c8 = \u2212 sin(L1) cos(\u03bb1) cos(\u03c812)\u2212 sin(\u03bb1) sin(\u03c812)i \u2212 (sin(L1) sin(\u03bb1) cos(\u03c812)\u2212 cos(\u03bb1) sin(\u03c812))j + cos(L1) cos(\u03c812)k (6) P2 = cos(dist12)P1 + sin(dist12)U\u03c8 (7) L2 = tan\u22121 ( P2\u221a 1\u2212 (P2)2 ) \u03bb2 = tan\u22121 ( P2 P2 ) (8) The third main function used for the lateral navigation algorithm is the calculation of the position coordinate values of the intersection of two flight routes. If the air vehicle is routed by parallel deviation from a flight route, the intersection point of the two flight legs must be calculated for lateral guidance. The intersection point of two flight legs is shown in Fig. 2. P0 is the intersection of the P1 \u2212 P2 route and P3 \u2212 P4 route. L1, \u03bb1 are the position coordinate values of P1, similarly L2, \u03bb2 are the coordinates of P2, L3, \u03bb3 are the position coordinate values of P3, and L4, \u03bb4 are the position coordinate values of P4. The longitude and latitude values of the intersection point P0 is L0, \u03bb0. \u03b7P1P2 = P1 \u00d7 P2, \u03b7P3P4 = P3 \u00d7 P4 (9) P01 = \u03b7P1P2 \u00d7 \u03b7P3P4 |\u03b7P1P2 \u00d7 \u03b7P3P4| P02 = \u03b7P3P4 \u00d7 \u03b7P1P2 |\u03b7P3P4 \u00d7 \u03b7P1P2| (10) In order to find the coordinate values of the intersection point, initially the normal vectors of the two surfaces are calculated. These two surfaces intersect at two waypoints. These two points are calculated by the formula shown above. By using Pnk(i), the latitude and the longitude coordinate values of the intersection point are calculated. Pnk(i) shows the ith element of the Pnk vector. L0 = tan\u22121 ( P01(3), \u221a 1\u2212 (P01(3)) 2 ) (11) \u03bb0 = tan\u22121 (P01(2), P01(1)) (12) If a parallel deviation from the flight route is required, the pitch and bank angle correction commands are calculated for fitting the desired flight path. Moreover, when the flight route switching waypoints are reached, these correction angles are calculated again in for introducing the next flight leg. [8] In order to calculate the lateral guidance command, firstly the projection waypoint of the current position of the air vehicle on the desired route is found. KTCP distance shown in Fig. 3, is the distance between the projection waypoint of the current position of the air vehicle and the point that the air vehicle will pass on the route. The position coordinate values of the point that the air vehicle will pass on the flight leg is found by the formula 7. The course angle between the current position vector of the air vehicle and the point RP is found by the formula 4. This course angle is called the desired track angle. ATD is the along track distance, the distance between the vertical projection waypoint of the current position of the air vehicle and the waypoint to be reached. Xtd is the vertical distance to the flight route. \u03c8track = tan\u22121 ( VGS E VGS N ) (13) The angle between the ground speed components is found. This angle is called the track angle. Then the track angle error is calculated by subtracting track angle from the desired track. \u03c8error = \u03c8desired \u2212 \u03c8track (14) By using the track angle error, lateral rotation speed command is found. Then finally by using the formula 16, the lateral guidance bank angle command is calculated. The lateral bank angle command provides holding the air vehicle in the desired route. \u03c8\u0307commanded = (2\u03c0fd lat)\u03c8error (15) \u03c6commanded = tan\u22121 ( (V 2 GS N + V 2 GS E) 1/2.\u03c8\u0307commanded g ) (16)"
        ]
    },
    {
        "image_filename": "designv11_64_0003535_iwc.2016.8068367-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003535_iwc.2016.8068367-Figure1-1.png",
        "caption": "Figure 1. Classification of wheel RCF by Deuce [1]",
        "texts": [
            "eywords: High speed wheels, rolling contact fatigue, monitoring tests, crack initiation; re-profiling 1 Introduction Nowadays rolling contact fatigue (RCF) has become a prevalent problem on all types of railway systems. It is believed that RCF comes into being as accumulated plastic flow exceeds the fatigue or ductility limits of surface material. According to its position, Deuce [1] classified continuously distributed wheel RCF into 4 categories, occurring in zones 1-4 illustrated in Figure 1, respectively. Zone 1 is located outside the nominal rolling circle (NRC, 70 mm away from the inside of wheel rim), so that the RCF in zone 1 is occasionally referred to as field side RCF. Among the 4 categories, the RCF in zone 1 is the most common type in practice [1, 2, 3]. This study focuses on a RCF problem of Chinese high-speed Electrical Multiple Unit (EMU) trains, see Figure 2 for examples. Obviously, the observed RCF is of the type in zone 1. Intensive investigations on RCF in the lasted decades have recognized that effective counter measures against wheel/rail RCF include improved steels, profile optimization, re-profiling/grinding, and friction management"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001519_ijnt.2014.059831-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001519_ijnt.2014.059831-Figure1-1.png",
        "caption": "Figure 1 Schematic illustration of the preparation of RGO/SF-based composite film (see online version for colours)",
        "texts": [
            " The films were heat-treated at 450\u00b0C for 30 min to remove the alcoholic solvent. To manufacture the DSSC, the prepared porous TiO2 thin film electrode was immersed in N719 dye solution at room temperature for 24 h, rinsed with anhydrous ethanol and dried. The graphene and silk fibroin based composite film were prepared by spin-coating once as the counter electrode was placed over the dye-adsorbed TiO2 electrode. The redox electrolyte consisted of 0.50 M KI, 0.05 M I2 and 0.5 M acetonitrile as solvent. Figure 1 shows the preparation process of GSF films. Due to wetting surface of GO sheets and dewetting surface of Si (silicon) wafer, it is difficult to spin-coat GO solution onto Si wafer. To combat this, the Si wafer was treated by dipping in APTES (3-aminopropyltriethoxysilane) solution (12 \u00b5L APTES in 20 mL water) for 20 min, forming a wetting surface to increase the coating rate of the GO solution. Completing the process of GO solution-dropping, spin-coating, chemical reduction treatment, silk fibroin solution-dropping, air drying, and peeling (for details, see Materials and Methods) result in GSF-based composite films(GSF films)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003908_1.1698360-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003908_1.1698360-Figure5-1.png",
        "caption": "FIG. 5. Representation of the elastic curve showing the contraction of the beams with increasing torsion angle.",
        "texts": [
            " Subscript t indicates bearings under tensile stress; subscript c indicates bearings under compression stress. For equal changes in temperature of both bearings the above condition reads (b.2 'a./ (S.) 0.5) / (b,2. a,/tS,) 0.5) = 0.89. (23) To keep the torque of such a double bearing zero independent of the angle of rotation, the relative variations in length U/L of the beams of both bearings due to a rotation also have to be equal otherwise Eq. (23) no longer applies. The total 300 change of length of a single beam due to a rotation through the angle cp (see Fig. 5) can be considered to consist of : (a) A lengthening caused by the curvature of the beam. (b) A contracting 0 caused by the motion of point II on the circle with the radius R 2\u2022 The sum of both changes results in a contracting which is, in case of the compression bearing, (24) Choosing for the tension bearing the ratio L/\"A =3.15, the relative change of the beam length equals the expression of Eq. (24). This is the reason for the choice of that ratio. Now it is comparatively easy to calculate for both bearings the restoring torque resulting from any rotation cpo For the compression bearing results are as follows: The torque of the tension bearing is given by llMot = 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.54-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.54-1.png",
        "caption": "Fig. 2.54 Three-element truss structurewith different external loading: a force boundary condition; b displacement boundary condition",
        "texts": [
            " Use a single (a) linear and (b) quadratic rod element to determine \u2022 the reduced system of equations, \u2022 the elongation of the rod at x = L , \u2022 simplify your result for the special case k = 0, \u2022 simplify your result for the special case E A = 0. \u2022 Compare the finite element solution with the analytical solution for the case k = 3, E A = 1 and L = 1. 2.5 Supplementary Problems 85 Fig. 2.53 Rod with elastic embedding loaded by a single force 2.35 Plane truss structure arranged in a square Given is the two-dimensional truss structure as shown in Fig. 2.54. The three truss elements have the same cross-sectional area A and Young\u2019s modulus E . The length of each element can be taken from the dimensions given in the figure. The structure is loaded by (a) a horizontal force F at node 2, (b) a prescribed horizontal displacement u at node 2. Determine for both cases \u2022 the global system of equations, \u2022 the reduced system of equations, \u2022 all nodal displacements, \u2022 all reaction forces, \u2022 the force in each rod. 2.36 Plane truss structure arranged in a triangle Given is the two-dimensional truss structure as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002584_ijamechs.2015.074786-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002584_ijamechs.2015.074786-Figure1-1.png",
        "caption": "Figure 1 Underwater vehicle (robot) with a manipulator",
        "texts": [
            " Furthermore, its controller structure is indeed much simpler than those of the adaptive controllers in Taira et al. (2010, 2012). In our previous paper (Taira et al., 2014b), a robust controller with a fixed compensator was designed. However, the stability of the control system including the compensator is not guaranteed. On the other hand, the stability (the ultimate boundedness of a tracking error) is ensured for the controller proposed in this paper. Consider an underwater vehicle (robot) equipped with a NL link manipulator that has only revolute joints, as shown in Figure 1. The explanations of main symbols used in this paper are provided in Table 1. Moreover, without loss of generality, we assume that NL = NM. As in Antonelli (2003) and Yoerger et al. (1990), the mathematical models of a UVMS with thruster dynamics are expressed as ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( , , ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) M V x t J u t R Kz t M u t f u d g \u03c4 t z t AD v z t BD v \u03c4 t = \u23ab \u23aa \u23a1 \u23a4\u23aa+ + = \u23ac\u23a2 \u23a5 \u23a3 \u23a6\u23aa \u23aa=\u2212 + \u23ad \u03c6 \u03c6 \u03c6 \u03c6 \u03c6 (1) 1( ) diag(| ( )|, ,| ( ) |) ( ) ( ) ( ) V V V V N N N N D v v t v t R z t D v v t R \u00d7= \u2208 \u23ab \u23ac = \u2208 \u23ad \u2026 (2) where the notation 1diag( , , )na a\u2026 represents the diagonal matrix whose non-zero elements are 1a to "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002189_phm.2014.6988158-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002189_phm.2014.6988158-Figure4-1.png",
        "caption": "Fig. 4. The mounted location of every accelerometer",
        "texts": [
            " And between the two of them is connected by coupling, as shown in Fig. 2. The wear fault is implemented on one tooth of sun gear. Fig. 3 shows the sun gear after machine the fault and its original photo. Four accelerometers are mounted on the planetary gearbox casing by glue, wherein accelerometer 1 and 2 are mounted on the input side of the gearbox (1 is horizontal and 2 is vertical), accelerometer 3 is on the top of the casing and accelerometer 4 is fixed on the output side. The specific location of every accelerometer is as depicted in Fig. 4, and Fig. 5 is the structure of the planetary gearbox. The configuration parameters of planetary gearbox see Table I. Shaft 1, is driven by the motor at a speed of 1200 rpm. And the sun gear is machined on it. The magnetic powder brake provide three different kinds of loads, they are 0 Nm, 0.6 Nm and 1.2 Nm. The characteristic frequencies of the system are calculated in Table II. The sampling frequency of this experimental system is 20K Hz, and 240000 points of data are recorded for every signal"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002948_gt2016-57458-Figure19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002948_gt2016-57458-Figure19-1.png",
        "caption": "Figure 19 Assembly procedure of optimized SLM produced front panel.",
        "texts": [
            " As a single part was produced to be tested in the engine, changes to the existing process were limited with respect to manufacturing and weld seam location. The brazed front panel is attached to the cast burner via TIG welding. The welding process was not changed; instead the SLM front panel was adapted to allow for the same production method to be applied. By keeping the overall design similar to the brazed variant, limited to no modification of production tooling was needed. The assembly procedure of the SMZ reheat burner equipped with a SLM front panel is described as (see Figure 19): 1) Cast of burner with front panel 2) Removal of front panel, machining of burner hooks and preparation for welding 3) SLM Damping Front Panel 4) Welding of front panel 5) Final machining of burner interfaces (slots, seals\u2026) When combining SLM components with existing parts the behavior is largely unknown. The majority of the experience is with single pass welding, where the wall thickness is thin enough to allow full penetration. In this setup multiple passes are needed as the wall thickness and geometry do not allow single pass welding"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000821_amm.756.85-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000821_amm.756.85-Figure3-1.png",
        "caption": "Fig. 3 - Operational scheme of t triangulation laser sensor",
        "texts": [],
        "surrounding_texts": [
            "Hardening with surface plastic deformation (SPD) is sufficiently well-known method to improve the operational characteristics of the machine components. One of the main technological parameters of SPD is the load, applied at plastic deformation of hardened surface; its nature and value in large measure determine depth and hardening degree, as well as the requirements to processing equipment, tool and machining attachment. Therefore, in the methods of SPD it is very important to accurately and effectively control the deforming load. By the nature of deforming impact SPD methods shall be divided into static and shock ones. Usage of dynamic load is energetically more cost-efficiently, but the one of the reasons that restrict the extended application of shock methods of SPD is the complexity of exercising of control over shocks energy, under the effect of which the plastic deformation of hardened material is occurred that often does not allow with sufficient accuracy to predict the results of hardening. According to conducted researches it is found that extensive reserves of improving of shocks energy use at plastic deformation are in creation of the most effective for this material form shock pulse due to deformation waves, propagating in shock system upon shock [1-3]. However, the efficiency of wave processes usage is greatly reduced or minimized, if it is impossible to accurately determine and regulate the kinetic energy of the shocks. This problem is most current one for SPD methods, designed to obtain a large depth of the hardened surface layer, where usually, as equipment shock mechanisms are used that initially are not intended for plastic deformation, such as hammers for crushing of solid materials during construction or mining operations. Therefore, before usage of such shock mechanisms for SPD their adaptation is needed that is serves for structural change of shock system to ensure the usage of wave processes during elastic-plastic deformation [4, 5]. At that it is necessary to conduct researches to determine the kinetic shocks energy of the impact and to ensure its precise regulation. Determination of the kinetic energy for hammers with a known mass of the hammer head shall be come to the determination of the shock velocity. To measure the velocity of the hammer head at its movement during speedup the inductive speed sensors, which transform translational motion of the hammer head to electrical signal, may be used. Such scheme was used for shock mechanism to harden SPD - hydraulic pulse generator (fig. 1), designed on the basis of known structures of hydraulic hammers for mining operations [4]. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications Ltd, www.scientific.net. (#523754121, Link\u00f6pings Universitetsbibliotek, Link\u00f6ping, Sweden-04/01/20,07:04:38) Technical characteristic of the hydraulic pulses generator Shock energy, J\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026.60...300 Adjustment range of shocks frequency, Hz\u2026\u2026..\u2026\u20263...40 Pressure of process fluid, \u041cP\u0430 \u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026.16 Flow of process fluid, l/min \u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026..\u2026\u202650 Maximum capacity, kW\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026.\u2026\u20268.5 Mass, kg\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026.\u2026\u2026 90 Overall dimensions, mm\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026220\u00d7 950\u00d7 295 Metal consumption, kg/J \u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026..\u2026\u20260.3 Efficiency\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026..\u2026\u20260.6 To determine the change of hammer head speed during motion at different time moments it was applied inductive speed sensor that converts reciprocating motion of the hammer head into an electrical signal. Inductive sensor is rod 11 that is coaxially fixed on workless end of the hammer head 1 and having the possibility of free reciprocating movement in tube 12 on which single-layer measuring coil 13 is wound that is connected to a storage oscilloscope 14 and polarizing coil 15 with independent power source 16 of 12 V. The length of the measuring coil 13 shall be greater than the operating stroke of the hammer head by factor of 1.5. During hammer head motion with rod 11 in the measuring coil 13, e.m.f. is induced that is proportional one to the velocity of the hammer head 1, and specially designed polarizing coil 15 provides flux closure, allowing to amplify the signal, supplied from the measuring coil to oscillograph 14. Oscillograph signal is recorded as oscillogram of voltage changes in time. To convert the voltage into the speed of the hammer head the calibration of variable reactance transducer was carried out. The tube 12 was chucked in cutter support of a lathe, and the end of the rod 11 was pivotally fixed on the disc, mounted in the machine spindle (representing crank mechanism), so that the spindle rotation was converted into linear motion of rod 11. On the basis of known spindle speed and geometric dimensions of the disk, the speed of the hammer head motion for different voltage, recorded on oscillograph was calculated. As a result of carried out experiments oscillograms, characterizing the change of speed of hammer head motion at various times during operation of the pulses generator (fig. 2). Disadvantage of application of speed variable reactance transducers is the necessity of their mounting on the hammer head of shock mechanism, but it is not always possible to implement. Method requires calibration and tare of used transducers. To get the chart of hammer head motion, the triangulation measurement method, using a laser sensor with a high sampling rate is sufficiently promising one. The main task, solved by triangulation sensor - is non-contact determination of the distance to the measurement object. Formerly, the method of laser measurement of position was used in geodesy and cartography. The application of laser sensors for measuring of moving objects was constrained by their slow response. At present time for usage the wide range of high-speed sensors is available. The operational principle of the laser sensor is based on the triangulation method of measuring of the distance to the object. Radiation of a semiconductor laser 1 is focused by optical objective 2 on the object 6. Radiation that is scattered on the object by optical objective 3 is collected on CCD-rule 4. Movement of investigated object from position 6 to position 6\u2019 causes corresponding movement of the image. Signal processor 5 calculates the distance to the object according to position of the light spot image on rule 4 and transmits the measured value to output of the sensor. Signal from the sensor can be transmitted to computer both directly through port: RS232, RS485 and via a digital oscillograph. 1 - Semiconductor laser 2, 3 - optical objective, 4 - CCD-rule, 5 - signal processor, 6,6` - investigated object. When machining of small-envelope parts when for hardening it is enough shock pulses energy, not exceeding 10 ... 50 J, we can effectively use scale electric hammers, intended for construction works (destruction of solid, compacted and frozen soils, for breaking of concrete surfacing, asphalt and masonry). The merit of usage for of RPE for SPD hardening of such devices is initially laid in design the independent regulation of energy and shocks frequency, small size, relatively low cost (hydraulic hammers), relatively (pneumatic hammers) high efficiency (40 ... 50%), lack of need to use additional devices of electric power supply of drive, such as oil pumping station for hydraulic hammers or compressor for pneumatic hammers. The operating principle of the electro-pneumatic hammer is in the following. Electric motor, mounted in the housing 1 (fig. 4) rotates the crank shaft 2. At that piston-rod 3, connected to the crank, transmits reciprocating motion to 4. During upstroke, in pneumatic chamber 5 of hammer head 6 below the piston the vacuum occurs. As a result of it, the hammer head starts to move up following piston, cocking for speedup. Then the piston, passing the top dead center, begins to move down against the hammer head, and the air, trapped between the hammer head and the piston begins to contract. Hammer head stops, and then starts moving down (operating stroke). At the end of the operating stroke the hammer head hits butt end of wave guide that has ability to move freely within axle-box 8. To determine the instantaneous coordinate of hammer head displacement by triangulation method, instead of wave guide with deforming tool, it is necessary to install waveguide 7, having a longitudinal opening for passage of the laser beam from triangulation laser sensor 9, installed within the power dog 10, disposed on fixed stand 11 of the unit for mechanical hardening. Direct laser beam 12 from the sensor through the longitudinal opening in the wave guide gets on the moving butt end of the hammer head that makes operational cycle, and the reflected beam 13 returns to the sensor, which produces 7700 measurements per second. Obtained coordinates are transmitted via Ethernet to PC. Technical specification of electric generator of mechanical impulses Maximal shocks energy, J\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026 25 Shocks frequency, Hz\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u202623.3 Press force, at least, N\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026 160 Power consumption, kW\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u20261.5 Useful voltage, V\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026... 220 Useful frequency, Hz\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026 50 Overall dimensions, mm\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026475\u00d7120\u00d7235 Weight (without cable and operating tool), kg\u2026 15 Metal consumption, kg/J \u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u20260.6 Efficiency\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u20260.4 As a result of performed tests, the indicator diagram of hammer head position (fig. 5), by which the momentary velocity of hammer head was defined, including the one that prior to shock \u03bdwas obtained. Motional energy of single shock is being defined by the following 2 2 hm A \u03bd\u22c5 = , Where, m is hammer head weight. Also, using the indicator diagram and movement speed graph it is possible to define the hammer head shocks frequency."
        ]
    },
    {
        "image_filename": "designv11_64_0002933_gt2016-56900-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002933_gt2016-56900-Figure1-1.png",
        "caption": "Figure 1. Seal structure of the AMB-HPS integrated seal",
        "texts": [
            " Inside the back-to-back compressors, larger axial space is occupied by the seal inside it, especially at the balance piston. There is large differential pressure on both sides of the centrifugal compressor\u2019s balance disc seal, which will make the cross coupling stiffness caused by the fluid rotation in the seal clearance more prominent. The cross coupling stiffness is responsible to reducing the effective damping of the seal and then generating the rotordynamic instability. An electromagnetic actuator presented in this paper is adopted to replace the original balance drum seal as shown in Fig. 1. Figure 1(a) shows the structure of original balance drum seal and Fig. 1(b) shows the improved seal arrangement. The stator of electromagnetic consists of silicon steel sheets, fixing bolts and retaining rings of stator-iron-core. Actuator with such a structure can apply electromagnetic force on the rotor of turbine machinery to control the vibration by changing the current in coils. Figure 1(c) shows the sealing body, which is inserted into the actuator, as shown in Fig. 1(c), to form a continuous surface on the inner surface of actuator. Electro-discharge machining or CNC drilling machining is used to produce the counterborings on the surface of inner circle. The structure of the seal is presented in Fig. 1(d). Figure 2 shows the 3-D graph of the structure of the backto-back centrifugal compressor\u2019s rotor, in which the actuator is installed. Here, the original balance disk is replaced by rotor of AMB-HPS. The labyrinth seal at this location is replaced by the 2 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89517/ on 02/04/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use stator of AMB-HPS. Figure 3(a) and (b) show the picture of the actuator with and without seal respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001202_argencon.2014.6868489-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001202_argencon.2014.6868489-Figure1-1.png",
        "caption": "Fig. 1. Simplified schematic of the plant used",
        "texts": [
            "975)e y x y x y x q\u03b1 =\u23a7 \u23aa =\u23a8 \u23aa = + + \u2212\u23a9 (2) In (1) and (2), the state variables x1, x2 and x3 denote drum pressure (kg/cm2), electric output (MW), and fluid density (kg/m3), respectively. The inputs u1, u2 and u3 are the valve positions for fuel flow, steam control, and feedwater flow, respectively. The output y3 is the drum water level (m), and \u03b1 and qe are steam quality and evaporation rate (kg/s), respectively and are given by: 3 1 3 1 2 1 1 3 (1 0.001538 )(0.8 25.6) (1.0394 0.0012304 ) (0.854 0.147) 45.59 2.514 2.096e x x x x q u x u u \u03b1 \u2212 \u2212 = \u2212 = \u2212 + \u2212 \u2212 (3) The Fig. 1 shows a simplified schematic of the plant used, The system presents the following physical restrictions in the control actions: 1 2 3 0.007 / 0.007 / 1/ 0.1/ 0.05 / 0.05 / 0 1 (1,2,3)i du du s s s s dt dt du s s u i dt \u2212 \u2264 \u2264 \u2212 \u2264 \u2264 \u2212 \u2264 \u2264 \u2264 \u2264 \u2200 \u2208 (4) Some typical operating points of the boiler-turbine model (1), given by [4], are tabulated in Table I. The literature ([4],[6],[7]) usually takes the operating point #4 as the nominal working point. For more details about the points of operation see [1], [4], [6], [7]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.9-1.png",
        "caption": "Fig. 6.9 Features of drilling cutter",
        "texts": [
            " The lower bound of Fdraw can be estimated from the z-stress around the sheet drawing Fdraw \u00bc \u03c0Ddraw\u03c3zH: Fdraw 1:155\u03c0\u03c3yieldDdrawH ln Do Ddraw \u00f06:17\u00de Typically, we may adopt a larger level for Fdraw: Fdraw \u00bc 1:5\u03c0\u03c3yieldDdrawH ln Do Ddraw \u00f06:18\u00de Machining is the process involving the removal of material by cutting the raw object into the expected product. Materials used in machining are widely covered, including metal, wood, plastic, and ceramic. There are three widely applied machining processes \u2013 drilling, turning, and milling. Drilling processes are machining operations for hole machining, which involve reaming and boring. Figure 6.9 shows the key features of drilling bits and gives a detailed description about them. Boring is used for enlarging an existing drilled hole; reaming is used for finishing an existing hole; and lapping is used for the surface finishing. Drilling processes can generate a significant material removal rate. They can achieve medium to high surface finish quality and good dimensional control of the holes they produce. By applying different configurations of drilling parameters, holes with different sizes and depths can be created in a variety of materials"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000990_amm.555.192-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000990_amm.555.192-Figure4-1.png",
        "caption": "Fig. 4. Programmed speed modification \u2013 time approach",
        "texts": [
            " = + (4) where has the direction of and its value is defined by decreasing function ( ) with = for = 0, has the direction of and its value is defined by increasing function ( ) with = 0 for = 0. In the second approach, named time approach, the closest point on the curve is defined by the programmed position for the current time. Like in the previous case the programmed speed at point is modified according Eq. 4, but the component is set along the direction . It means that in case the current position is behind the programmed one for the current time the value of programmed speed is increased (like for positions and in Fig. 4). In case the current position is ahead of the programmed one for the current time the value of programmed speed is decreased (like for position in Fig. 4). To build a simulation model the programmed trajectory is defined as a curve which consists of a set of Bezier segments. A j-th segment is based on four control points defined by third order equations. Coordinates ( , ) of curve points for two dimensional j-th segment are given by Eq. 5 and Eq. 6. ( ) = (1 \u2212 ) + 3 (1 \u2212 ) + 3 (1 \u2212 ) + (5) ( ) = (1 \u2212 ) + 3 (1 \u2212 ) + 3 (1 \u2212 ) + (6) where , and , are start and end points of the segment respectively, , and , define the curve tangency direction at start and end points respectively, is a parameter which varies from 0 to 1",
            "02[ ], 0.8[ ]), (0[ ], 0.8[ ]), - Bezier programmed trajectories for robot A (experiments a, b, c and d) are defined by points given in Table 1. Result trajectories for the above data and geometric method of calculation of programmed speed according to Fig. 3 are presented in Fig. 5 (in robot A coordinate system). Maximal distances between generated and programmed trajectories are given in two last columns of Table 1. Result trajectories for time method of calculation of programmed speed according to Fig. 4 are presented in Fig. 6. Machines-Tools, Mechanical Engineering and Human Motricity Fields Simulation 2. This simulation is intended to present influence of programmed robot velocities on the shape of generated trajectories. Initial and final positions are defined as in simulation 1. Programmed trajectories are defined by control points: - for robot A: (0[ ], 0[ ]), (0.15[ ], 0[ ]), (0.3[ ], \u22120.2[ ]), = (0.3[ ], \u2212 0.4[ ]), (0.3[ ], \u22120.6[ ]), (0.15[ ], \u22120.8[ ]), (0[ ], \u22120.8[ ]), - for robot B: (0[ ], 0[ ]), (0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003090_icelmach.2016.7732872-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003090_icelmach.2016.7732872-Figure1-1.png",
        "caption": "Fig. 1. Current superimposition variable flux reluctance machine.",
        "texts": [
            " The VFRM is composed of 2 sets of windings: armature windings for generating a rotating magnetic flux, and field windings for controlling the magnetic field intensity. Due to this, the torque constant of the VFRM can be controlled and its power band can be increased. However, since two separate sets of windings are required, the size of the motor is large and it becomes more complicated to manufacture. In order to solve these problems, a current superimposition variable flux reluctance machine with concentrated winding (CSVFRMCW) has been proposed (Fig.1). The machine requires only a single set of windings that can perform both armature and field winding functions simultaneously [17]-[19]. By using the single set of coils, the winding structure is simplified. It is easy to apply the distributed winding and sprit core to our proposed machine. In addition, a high productivity is expected. In this paper, we propose a current superimposition variable flux reluctance machine using distributed winding (CSVFRMDW). The structure, winding pattern and \u03a6 A"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure6-1.png",
        "caption": "Fig. 6 Solution of topology optimization problem \u2013 caliper (left) and upright (right).",
        "texts": [
            " The volumetric displacement of the caliper is given by the total volume variation in the hydraulic cylinders due to the deformation of the caliper when subjected to braking loads [8]. The volumetric displacement is somehow related to the pedal displacement that the driver feels when activating brakes since it is directly proportional to the amount of fluid that is pressed in the hydraulic cylinders. A penalization factor equal to 4 has been set for the optimization algorithm. Convergence to an optimal solution was reached after 36 iterations. The final structural layouts are shown in Fig. 6. The caliper exhibits an asymmetric structural layout that confirms the results obtained in other works performed by the authors [8]. The asymmetry is mainly due to the asymmetric boundary conditions (the caliper is fixed to the upright on one side while the other is free). Structural performances of the obtained solution are reported in Tab. 2 and compared with actual components. Tab. 2 Structural performances of the obtained solution and comparison with the actual component. Performance index Value Variation w"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001350_15325008.2014.943438-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001350_15325008.2014.943438-Figure1-1.png",
        "caption": "FIGURE 1. The parts of the EMA.",
        "texts": [
            " The iron yokes fix the structure, including upper iron yoke, lower iron yoke, and side iron yoke. The flux path is formed of soft magnetic material to reduce magnetic leakage. The movable axle center consists of axle center and push rod, made of soft magnetic material and stainless steel, respectively, and moves upwards when attracted by magnetic force. The copper sheathing and other fittings are made of non-magnetic material. Copper sheathing can eliminate the residual magnetism of the movable axle center [11]. The parts of the EMA used in this study are shown in Figure 1. In this study, 125 Vdc was applied to EMA. The trip coil excited a current to magnetize the movable axle center, which formed an electromagnet that generated a magnetic force. When the excited current gradually increased, it was able to overcome the weight of the movable axle center, which then moved upwards. The actuation of the EMA contained input, trip coil excitation, no-load motion of the trip coil axle center, axle center pushing trip trigger, main trip after tripping of trip trigger, and axle center returning to the home position after a trip action"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002935_gt2016-56508-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002935_gt2016-56508-Figure2-1.png",
        "caption": "Figure 2: Two-pad offset-preloaded AFB configuration",
        "texts": [
            " The high speed bench test, which is driven by an impulse turbine, is designed to spin the rotor up to 185krpm. Two-pad offset-preloaded air foil bearing (AFB) and sixpad air foil thrust bearing (AFTB) configurations are adopted in this work. Compared to single-pad AFB, the two-pad design inherits better stability at high speeds due to its anisotropic stiffness characteristics and very low cross coupled stiffness as explained later. The general design concept of the AFB, which is schematically depicted in Figure 2, can be found in [9]. The bearing parameters, as listed in Table 1, are adjusted in accordance to the shaft diameter and load capacity requirement. Readers can also refer to Figure 1 for further clarification of the bump dimensions. The designed hydrodynamic preload, which is the offset between the bearing sleeve center and the top foil curvature centers, is 60\u03bcm. The AFTB load capacity is not yet verified. However, the thrust foil bearing was successfully adapted to commercial 12kW micro gas turbines, and the bearing provides sufficient load capacity to withstand the thrust load induced from the impulse turbine"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001429_wcica.2014.7053298-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001429_wcica.2014.7053298-Figure1-1.png",
        "caption": "Fig. 1. Possible communication topologies for G = {G\u03c3 |\u03c3 = a,b,c,d}",
        "texts": [
            " The reason is that the structure of the leader\u2019s model r\u03071 = r2 is not the same as the structure of the ith follower\u2019s model (21). In this section, we assume that the dynamics for all agents are identical, and simulate two examples. The results show four agents reach state consensus through local communication using the algorithm (6) and (22), respectively. Example 1 (Consensus for second-order nonlinear models with switching topologies): These directed graphs, which all have a spanning tree, associated with G are shown in Fig. 1 for algorithm (6). The network switches every T = 1 second to the next state according to the state machine in Fig. 2. And the model for the ith agent is chosen as Fig. 2. Finite automaton with four states representing the discrete-states of a network with switching topologies. Let c1 = 10, and c2 = 20 in (6). The trajectories of agent\u2019s states are shown in Fig. 3 and in Fig. 4. These figures show that the two states of the systems can reach consensus asymptotically. Example 2 (Consensus for second-order nonlinear models with unknown parameters): The fixed directed graph associated with the graph G , is shown in Fig. 1(a), and the 1st agent has access to the leader. For simplicity, there exists one unknown parameter and let the model of the ith agent be \u03b7\u0307i = 2\u03b72 i +(1+\u03b72 i )\u03bei, \u03be\u0307i = 2\u03b7i\u03bei + ui 10+ cos(\u03b7i)+ sin(\u03bei) . where \u03b8 = 2 is the unknown parameter, and \u03c61 = \u03b72 i and \u03c62 = \u03b7i\u03bei are known nonlinear functions. The leader\u2019s model is selected as r1 = sin(t), r2 = cos(t). Fig. 5, Fig. 6, and Fig. 7, show the trajectories of \u03b7i(t), \u03bei(t), and \u03b8\u0302i(t) by the algorithm (22). Note that the state \u03b7i tracks the first state of the leader\u2019s dynamic, which means that \u03b7i reaches consensus, and Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002394_j.jestch.2016.05.003-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002394_j.jestch.2016.05.003-Figure1-1.png",
        "caption": "Fig. 1. Test rig (a) schematic. (b) pictorial diagram.",
        "texts": [
            " (16) as, Xn i\u00bc1 yixi1 \u00bc na0 Xn i\u00bc1 xi1 \u00fe a1 Xn i\u00bc1 xi1xi1 \u00fe a2 Xn i\u00bc1 xi2xi1 \u00fe a3 Xn i\u00bc1 xi3xi1 \u00fe a4 Xn i\u00bc1 xi4xi1 \u00fe a5 Xn i\u00bc1 xi5xi1 \u00f017\u00de Xn i\u00bc1 yixi2 \u00bc na0 Xn i\u00bc1 xi2 \u00fe a1 Xn i\u00bc1 xi1xi2 \u00fe a2 Xn i\u00bc1 xi2xi2 \u00fe a3 Xn i\u00bc1 xi3xi2 \u00fe a4 Xn i\u00bc1 xi4xi2 \u00fe a5 Xn i\u00bc1 xi5xi2 \u00f018\u00de Xn i\u00bc1 yixi3 \u00bc na0 Xn i\u00bc1 xi3 \u00fe a1 Xn i\u00bc1 xi1xi3 \u00fe a2 Xn i\u00bc1 xi2xi3 \u00fe a3 Xn i\u00bc1 xi3xi3 \u00fe a4 Xn i\u00bc1 xi4xi3 \u00fe a5 Xn i\u00bc1 xi5xi3 \u00f019\u00de Xn i\u00bc1 yixi4 \u00bc na0 Xn i\u00bc1 xi4 \u00fe a1 Xn i\u00bc1 xi1xi4 \u00fe a2 Xn i\u00bc1 xi2xi4 \u00fe a3 Xn i\u00bc1 xi3xi4 \u00fe a4 Xn i\u00bc1 xi4xi4 \u00fe a5 Xn i\u00bc1 xi5xi4 Xn i\u00bc1 yixi5 \u00bc na0 Xn i\u00bc1 xi5 \u00fe a1 Xn i\u00bc1 xi1xi5 \u00fe a2 Xn i\u00bc1 xi2xi5 \u00fe a3 Xn i\u00bc1 xi3xi5 \u00fe a4 Xn i\u00bc1 xi4xi5 \u00fe a5 Xn i\u00bc1 xi5xi5 \u00f020\u00de Eqs. (16)\u2013(20) are simultaneously solved for the unknowns. After substitution of unknowns in Eq. (14), represents the model for prediction of the vibration acceleration amplitude of the defective roller in rolling contact bearings. To verify the functional form of the developed model Eq. (13), we conducted experiments on the test setup shown in Fig. 1. The rotor is supported by rolling contact bearings at both the ends. At the drive end self-aligning spherical roller bearing is fitted. At the non-drive end, test bearings are fitted having seeded defects on rollers. The rotor is driven at multiple speed by coupling it to a DC motor using flexible coupling. Motor current frequency controller was used to obtain the different rotor speeds. The test Please cite this article in press as: I.M. Jamadar, D.P. Vakharia, An in-situ synth Tech., Int. J. (2016), http://dx",
            " It is important to note that the vibration response as defied by Eq. (13) depends only on 5 dimensionless terms instead of the original 28 terms involved in the problem which are systematically reduced by applying partial correlation analysis of the problem using the matrix method of dimensional analysis. A set of systematically planned experimental data set listed in Table 5 has been used for obtaining the relationship developed in Eq. (14) after conducting experiments on the developed test setup shown in Fig. 1. The validation experiment is a final step in verifying the conclusions from the previous round of experimentation. In order to validate the results obtained from Eq. (14), confirmation experiments were conducted for obtaining the response characteristics in which a selected number of tests are run under specified conditions as listed in Table 7. The specifications of the test bearing for solution of the Eq. (14) are taken from Table 8. These bearings are damaged by inducing artificial spalls of different sizes to the rollers using electric discharge machining as shown in Fig",
            " 4 at 15 Hz. Please cite this article in press as: I.M. Jamadar, D.P. Vakharia, An in-situ synthesized model for detection of defective roller in rolling bearings, Eng. Sci. Tech., Int. J. (2016), http://dx.doi.org/10.1016/j.jestch.2016.05.003 For the validation test 2, the bearing 30305C was replaced by 32205BJ2/Q bearing. For the test 2, rotor speed was increased to 1300 rpm. Fig. 5 shows the experimental frequency spectrum obtained from the housing of the test bearing on the test setup shown in Fig. 1. As it is observed from the spectrum, due to the increased rotor speed, shaft unbalance and the level of roller damage, the peak amplitude at the roller defect frequency has increased to a high value as compared to test 1 as listed in Table 10. A significant peak of 0.835 m/s2 is observed at the 365 Hz which is nearly equal to the 3rd harmonics of the theoretical roller defect frequency of 123.3 Hz. The other dominating peaks in the spectrum are at the combinations of the shaft and the cage frequency which are 10fs 3fFTF and 15fs 6fFTF i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.22-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.22-1.png",
        "caption": "Figure 9.22 Vector diagram according to the cos\u03c6= 1 control at a rated torque and a rated frequency. Usph= 169 V, EPMph= 186 V, Is= 132 A, cos\u03c6= 1, and \u03b4s= 30.3\u00b0. The d- and q-axis currents are 0.57 and 0.99, respectively. Te= 0.92 pu.",
        "texts": [
            "4\u00b0 \u03c8 s 1.01 The power factor cos \u03c6 is calculated from the base of phase angle \u03c6, which is now identical to the load angle, because the current space vector is perpendicular to the flux linkage \u03c8PM and the voltage space vector is perpendicular to \u03c8s. Referring to the space-vector diagram, the power factor is cos \u03c6 0.91 The electromagnetic pu power and torque are as follows. Pe usiscos \u03c6 1 1 0.91 0.91 pu usiscos \u03c6 1.0 1 0.91 Te 0.91 pu \u03c9s 1 The same motor can be driven using a unity power factor. Figure 9.22 shows the resulting vector diagram. The current vector points in the negative d-axis direction resulting in smaller stator flux linkage, because of the armature reaction. Moreover, the voltage level is lower. The figure clearly indicates that to obtain an appropriate stator current at the rated torque, the flux linkage produced by the PMs should be higher. Here, the supply voltage is low, and as a result, the actual current demand exceeds the rated value by a considerable margin. There is plenty of voltage reserve"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure14.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure14.9-1.png",
        "caption": "Fig. 14.9 Asymmetrical load characteristic. a Internal resistance via its voltage. b Current of the internal resistance via its voltage and transfer characteristic. c Current via the internal conductivity",
        "texts": [
            "16) is the particular case of Rapp\u2019s model of a solid-state microwave power amplifier [8, 19]. If to express the current I through the internal conductivity Yi = 1/Ri, yi OC = 1/ri OC, the circle equation turns out Yi yOCi 2 \u00fe I AISC 2 \u00bc 1: \u00f014:17\u00de The plot of this circle is shown in Fig. 14.8c. 14.5 Asymmetrical Load Characteristics We may introduce asymmetrical characteristics of a common view too. For example, a transfer characteristic of MOSFET transistor may considerably be differing from a symmetric curve. Case 1 We consider a new position of the known hyperbola in Fig. 14.9a. Let the asymptotes 0y, 0x form a rectangular coordinate system y0x, which turned on an angle \u03b1 concerning the initial system Ri 0 Vi. Therefore, the equation of the hyperbola has the view 14.4 Symmetrical Load Characteristic \u2026 401 On the other hand, for the coordinates of point M, the following orthogonal transformation is known [10] x y \" # \u00bc cos a sin k sin a cos a \" # Vi Ri \" # : \u00f014:19\u00de Then, expression (14.18) obtains the view cos a sin a R2 i \u00fe cos2 a sin2 a RiVi cos a sin a V2 i k \u00bc 0: This view corresponds to the quadratic form [10] a11R 2 i \u00fe 2a12RiVi \u00fe a22V 2 i \u00fe a33 \u00bc 0: \u00f014:20\u00de Hereinafter, dimensions of the values are not used. Using transformation of the variable Ri by I, we get a quadratic equation I2 \u00fe 2 a12V2 i a22V2 i \u00fe a33 I\u00fe a11V2 i a22V2 i \u00fe a33 \u00bc 0: \u00f014:21\u00de The plot of this expression has the typical view in Fig. 14.9b with different values of the maximum currents IM + , IM \u2212 . If to express the current I through the internal conductivity Yi, the general circle equation turns out a11 \u00fe 2a12I\u00fe a22I 2 \u00fe a33Y 2 i \u00bc 0: From here, we get the explicit circle equation k cos a sin a Y 2 i \u00fe I IC 2\u00bc 1\u00fe IC 2 ; \u00f014:22\u00de where IC \u00bc cos2 a sin2 a 2 cos a sin a \u00bc 1 2 1 tga tga : The plot of this circle is shown in Fig. 14.9c. The maximum currents IM + , IM \u2212 correspond to Yi = 0. Therefore, I \u00feM \u00bc 1=tga; I M \u00bc tga: In turn, the current IC conforms to the center of the segment IM + IM \u2212 and determines the maximum value YiM. 402 14 Quasi-resonant Voltage Converter with Self-Limitation \u2026 If asymmetrical curve (14.21) is used for approximation of the transfer characteristic ID(VGS) of MOSFET transistor, it is necessary, in the initial coordinate system ID 0GS VGS, to restore the coordinate system I 0 Vi. To do this, we may use the following property of the tangent line into the cusp ID C, VGS C of the asymmetrical characteristic; this line passes through the origin of the coordinate system I 0 Vi"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001038_ipec.2014.6869699-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001038_ipec.2014.6869699-Figure7-1.png",
        "caption": "Fig. 7 Textile machine configuration",
        "texts": [
            " Starting of the machine under heavy load was possible. The response of the speed acceleration and deceleration were achieved within O.2s. Fig. 6 shows the waveform in position control mode, where the position varies along with the command under the heavy load of about 80% of the rated torque. The performance of the servo-locked position control was also satisfied to suppress the movement against the rebound torque caused in injection mode. The specification of the tested IPMSM is shown in Table 2. 3.3 Textile Machine Fig.7 shows the standard structure of the textile machine. Textile machinery market is an important market, which accounts for about 10% of the inverter annual sales in Japan. In recent years, the domestic market has become a difficult situation due to reduced production of synthetic fiber, including acrylic staple fiber. However, 80% of the production volume is for export, 70% of them, is for China. Future, the rise of India and Pakistan market is also attracting attention, expansion of the global market is expected"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure4.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure4.4-1.png",
        "caption": "Fig. 4.4 Superposition of the Bernoulli beam (a) and the shear deformation (b) to the Timoshenko beam (c) in the x\u2013z plane. Note that the deformation is exaggerated for better illustration",
        "texts": [
            " It is obvious that the equivalent constant shear stress can alter along the center line of the beam, in case the shear force along the center line of the beam changes. The attribute \u2018constant\u2019 thus just refers to the cross-sectional area at location x and the equivalent constant shear stress is therefore in general a function of the coordinate of length for the Timoshenko beam: \u03c4xz = \u03c4xz(x). (4.15) The so-called Timoshenko beam can be generated by superposing a shear deformation on a Bernoulli beam according to Fig. 4.4. One can see that the Bernoulli hypothesis is partly no longer fulfilled for the Timoshenko beam: Plane cross sections remain plane after the deformation. However, a cross section which stood at right angles on the beam axis before the deformation is not at right angles on the beam axis after the deformation. If the demand for planeness of the cross sections is also given up, one reaches theories of higher-order [31, 40, 41], at which, for example, a parabolic course of the shear strain and stress in the displacement field are considered, see Fig",
            " Note that the deformation is exaggerated for better illustration Following an equivalent procedure as in Sect. 3.2.1, the corresponding relationships are obtained: sin \u03c6y = ux z \u2248 \u03c6y or ux = +z\u03c6y, (4.16) wherefrom, via the general relation for the strain, meaning \u03b5x = dux/dx , the kinematics relation results through differentiation with respect to the x-coordinate: \u03b5x = +z d\u03c6y dx . (4.17) Note that \u03c6y \u2192 \u03d5y = \u2212 duz dx results from neglecting the shear deformation and a relation according to Eq. (3.16) results as a special case. Furthermore, the following relation between the angles can be derived from Fig. 4.4c \u03c6y = \u03d5y + \u03b3xz = \u2212duz dx + \u03b3xz, (4.18) which complements the set of the kinematics relations. It needs to be remarked that at this point the so-called bending line was considered. Therefore, the displacement field uz is only a function of one variable: uz = uz(x). 4.2.2 Equilibrium The derivation of the equilibrium condition for the Timoshenko beam is identical with the derivation for the Bernoulli beam according to Sect. 3.2.3: 4.2 Derivation of the Governing Differential Equation 195 dQz(x) dx = \u2212qz(x), (4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003955_ee.1933.6430720-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003955_ee.1933.6430720-Figure1-1.png",
        "caption": "Fig. 1. (Left) Flow of energy relations as expressed by eq 4 for resistance and inductance in",
        "texts": [
            " The rate at which energy is stored into the system or the work done by the force in producing motion against the inertia of the system is ~^^^2 \u00b7\u0302 The maximum inflow of stored energy therefore occurs when the stored energy is l / 2 its maximum and increasing, and therefore when the rate of dissipation is 1 / 2 its maximum value and increasing, and the maximum outflow of stored energy occurs at the same point of the stored energy cycle when it is decreasing, that is at the same point of the decreasing dissipation cycle. The cycle of inflow of stored energy is therefore in phase advance of the cycle of dissipation or power inflow by a right angle. These cyclic flow of energy relations are shown in Fig. 1, as expressed by eq 4 and in Fig. 3 for those who are happier when dealing with electromotive forces and currents these relations are shown as expressed by eq 5. For a circuit having potential energy, the stored energy is V V 2 C where q is the change at any instant and C is the capacity. The inflow of energy into the system is given by Ri2 + ( 6 ) (7) In the dynamical analogy q is the coordinate of the motion, i is the velocity. Starting with maximum velocity at time zero, the spring (supposed to be linear) will have reached 1 / 2 its maximum deflection and the maximum stored energy will occur when the velocity becomes zero, that is, when Ri2 is zero"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000419_978-3-642-33509-9_22-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000419_978-3-642-33509-9_22-Figure1-1.png",
        "caption": "Fig. 1. Omni-directinal mobile robot",
        "texts": [
            " We use an omni-directional robot with four omni-wheels and DC motors. The robot can move to different omni-direction by changing the combination of output levels to motors. Basically, the action outputs of the robot are direct forward movement and rotation at the same position to avoid the slip appeared as noise in SLAM. Furthermore, the robot changes the moving direction only when the robot conducts obstacle avoidance. In addition, we use a laser range finder (LRF, URG04-LN) for SLAM and self-localization. Figure 1 shows the omni-directional mobile robot. Figure 2 shows an example of multiple mobile robots. Sometimes, a mobile robot can enter the sensing range of other mobile robots (Fig.2 (a)). As a result, it is very difficult for the mobile robot to perform self-localization using a shared grid map (Fig.2 (b), because other robots appeared in the sensing range can be considered as unknown objects. In the proposed method, we use the occupancy grid mapping [14, 15]. Figure 3 shows the concept of the occupancy grid map"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003072_sbr.lars.robocontrol.2014.12-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003072_sbr.lars.robocontrol.2014.12-Figure3-1.png",
        "caption": "Figure 3: Robot Model Diagram",
        "texts": [],
        "surrounding_texts": [
            "As above mentioned, the NMPFC was implemented in a distributed fashion. For a team of robots to converge into a formation, the NMPFC\u2019s cost functions of each robot must be coupled. The above mentioned coupling occurs when the teammates\u2019 states are used in the cost function of each robot\u2019s controller to penalize the geometry or the deviation from the desired objective. In other words, the actions performed by each robot affects every other teammate. The NMPFC can be divided into two sub-blocks. The first, the Optimizer, uses an online numeric minimization method to optimize the cost function and generate signals of optimal control. The resilient propagation (RPROP) method is used here, which guaranties quick convergence [9]. The second sub-block, here called Predictor, performs the state evolution of the robot itself, the teammates and the target based on pre-defined models. Each robot keeps the formation state (pose and speed of the robots in formation, and position and speed of any target that should be followed), updating them in each control loop. This information is received by the controller of each robot in the formation which in turn creates the formation geometry where the actions of each robot affect the other teammates (as demonstrated in Fig. 2). After receiving the states of the robot and information of teammates and target, the controller\u2019s optimizer sub-block provides the control input, which then predicts the formation state evolution and provides a cost value to the optimizer in accordance with U\u0302(k + i|k) with i = 0...Nc\u2212 1, which is the output control signal from the optimizer sent to the predictor, in a limited control horizon Nc, to the predictor sub-block, which then predicts the formation state evolution P\u0302(k+ i|k) with i = 1...Np for Np steps (prediction horizons), which is the response of the predictor block to each U\u0302(k+ i|k) and provides a cost value to the optimizer. The iterative minimization process is repeated in cyclic fashion. Fig. 2 illustrates the structure of the NMPFC used in this work and presented by [3], where U(k|k) = U(k) = [ vre f (k) vnre f (k) wre f (k) ]T is the output control signal in the first prediction step N1. Here, at an instant k, robot 1 (R1) sends its pose PR1 (k) =[ xR1 (k) yR1 (k) \u03b8R1 (k) ]T to the NMPFC. Further- more, the NMPFC also receives the other robots\u2019 poses [PR2 (k)...PRN (k)], the position of the target t in the world frame wPt(k) = [ wxt(k) wyt(k) ]T and the velocity of the target t in the world frame wVt(k) =[ wvxt(k) wvyt(k) ]T . To achieve convergence in the formation, and hence cost function minimization, the NMPFC\u2019s predictor sub-block produces the evolution of the formation\u2019s behavior, as well as the behavior of the target\u2019s merged state and covariance matrix which is used by the NMPFC\u2019s optimizer and predictor sub-blocks for the cyclic minimization process. After processing the control calculations, the NMPFC sends the desired control output back to the robot (controller\u2019s reference velocities)."
        ]
    },
    {
        "image_filename": "designv11_64_0001862_065018-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001862_065018-Figure6-1.png",
        "caption": "Figure 6. (a) Ball trajectory measured for the 47.74 mm diameter billiard ball. (b) Measured ball velocity versus time (dots) and a linear fit to the data (solid line).",
        "texts": [],
        "surrounding_texts": [
            "Typical results for three of the balls are shown in figures 4, 5 and 6. Each ball spiralled inwards, as predicted, although the centre of the spiral path drifted slowly towards the upper left corner of the granite block or to the lower right corner depending on the direction of motion of the ball. The effect was due to a very slight tilt of the block, meaning that it was not exactly horizontal. The effect was not significant in terms of ball speed measurements since (a) the average radial drift speed away from the centre of the block was only about 0.003 m s\u22121 at most, and (b) no periodic increase or decrease in ball speed during each orbit was detected. Instead, the ball speed decreased linearly with time for at least the first 10 s, and for 30 s for the larger steel balls. There is a larger scatter in the velocity data for the billiard ball since the x, y coordinates were recorded at intervals of 0.05 s rather than 0.1 s, and since the position of the centre of mass was more difficult to determine accurately due to the larger ball diameter. Nevertheless, the acceleration of the ball was quite reproducible, giving 0.000 35 000 003rm =  averaged over four trials. For the smallest steel ball, 0.000 84 0.000 04rm =  averaged over three trials. For the 30 mm steel ball, 0.000 19 0.000 04rm =  averaged over four trials. For the 50.8 mm steel ball, 0.000 086 0.000 002rm =  averaged over three trials. As shown in figure 5(b), a better fit to the velocity data for the smallest ball was obtained with a quadratic rather than a linear fit, indicating that rm decreased slightly as the velocity decreased. The effect is more noticeable with the small ball since there was a larger change in ball speed. As a check on the theoretical prediction concerning the radius of curvature, the radius was measured from figure 4(a) by halving the diameter of each orbit. The experimental result is shown in figure 7, together with the theoretical prediction given by equation (5) and the measured ball velocity fit. For the ball in figure 4, MgdR I 0.1780 = m s\u22122. The agreement is very good, indicating that the gradual drift in the centre of curvature did not have a significant effect on the radius of curvature itself."
        ]
    },
    {
        "image_filename": "designv11_64_0000859_1.4027130-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000859_1.4027130-Figure15-1.png",
        "caption": "Fig. 15 A double-row trochoidal gear (Type C)",
        "texts": [
            " This means that increasing the contact ratio e due to an increase of the number of teeth in a singlerow trochoidal gear decreases the strength of the teeth and rollers. To overcome this problem, as a new transmission error reduction method, a double-row trachoidal gear (type C gear) was presented, Journal of Tribology JULY 2014, Vol. 136 / 031101-7 Downloaded From: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jotre9/930316/ on 04/13/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use as shown in Fig. 15 and its transmission error was examined by the experiments and MBA. 5.1 Experiments of a Double-Row Trochoidal Gear. For the experiments, a double-row trochoidal gear (type C gear) was made. The type C gear consists of a double-row roller gear and a double-row cam gear. In the type C gear, the locations of the rollers and teeth in one row were shifted to half of the pitch of the rollers and teeth, respectively. Because of this, the contact ratio e of the type C gear was two times the contact ratio per row"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure5.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure5.3-1.png",
        "caption": "Fig. 5.3 (a) An intrinsic crack-tip in material, (b) partially reversible dislocation motion, (c) deformation-induced twinning, and (d) degradation of bridging zones",
        "texts": [
            " Regardless, some intermetallics are lightweight and heat-resistant, such as gamma-based titanium aluminides (TiAl) and niobium aluminides (Nb3Al), which are of commercial interest for replacing heavier superalloys. The fatigue damage characteristics of intermetallics lie somewhere between metals and ceramics. Fatigue crack growth occurs as a result of intrinsic crack-tip processes (partially reversible dislocation motion, deformation-induced twinning, and crack-tip/environmental interactions) and crack wake processes (degradation of bridging zones), as shown schematically in Fig. 5.3. Stable fatigue crack growth occurs in intermetallics due to stress-induced martensitic transformations. The slowest fatigue crack growth rates are typically observed in intermetallics in which crack-tip dislocation processes are predominant. However, most intermetallics crack more quickly than metals and alloys. Instead of considering the time with loading, a more precise quantification is the number of loading cycles, Nlc. The relative contributions from each of these variables may be estimated from the modified two-parameter Paris equation, which is given by 124 5 Ceramics daf=dNk \u00bc CKE mTmax n \u00f05:7\u00de where n is the exponent of Tmax, which has a value typically close to 1, Tmax is approximately equal to the fracture toughness of the material Tf; and m is the exponent of the effective stress intensity factor KE"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000866_iros.2014.6943044-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000866_iros.2014.6943044-Figure2-1.png",
        "caption": "Fig. 2. A Cart consisting of j-Caster Units",
        "texts": [
            " Incidentally, since the proposed estimation method utilizes only the velocity information of each caster unit, it can be applied not only to our passive systems but also to activeactuated systems. The mobile robot platform using two double-wheel caster units with servo brakes is presented in Fig.1(a). The caster unit is installed with three encoders attached to two wheel shafts and a pivot shaft, as shown in Fig. 1(b). These encoders provide velocity information of the caster units. We first assume a rigid moving base with firmly attached caster units. The coordinate systems of j-caster units are illustrated in Fig. 2. One of the caster units, labeled the k-th caster unit, is selected as a representative unit. We consider the relationship between the i-th caster unit, and the representative, k-th caster unit. In Fig. 2, \u03a3i is the i-th coordinate system (where i = 1, . . . , j) originated at the pivot shaft of the i-th caster unit, and \u03a3k (k 6= i) is the coordinate system of the k-th caster unit. The attachment point of the i-th caster unit is defined as the origin of the coordinate system \u03a3i. The problem is to estimate the position and orientation of the coordinate system set on the attachment point of the i-th caster unit, relative to the equivalent coordinate system of the k-th caster unit. The translational and angular velocities of the attachment point coordinate systems are determined by   cx\u0307i cy\u0307i c\u03b1\u0307i   =   ri 2 ri 2 0 risi Ti \u2212 risi Ti 0 ri Ti \u2212 ri Ti 1     \u03c6\u0307ir \u03c6\u0307il \u03c6\u0307si   , (1) where ri, si, and Ti denote the radius of each wheel, the offset length, and the distance between two wheels, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003857_0954406215589843-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003857_0954406215589843-Figure2-1.png",
        "caption": "Figure 2. Position and motion of rigid pad model relative to the inertial frame at the bearing centre.",
        "texts": [
            " The aerodynamics of the air inside the thin film between pads and rotor are approximated sufficiently well by the solution of the Reynolds equation for compressible fluids. With these assumptions the dynamical interaction between rotor and bearing pads is only through aerodynamic forces and contact. For both, the evaluation requires tracking of the relative position and velocity of rotor and pads, that, because of the rigid body assumptions, are fully described by the small set of degrees of freedom (DOF). Figure 2 shows an example of how the motion of the compliant tilting pad can be described with three DOF of the pivot-point: Frame e is an inertial frame at the bearing centre Cb. Frame es has fixed orientations, but slides along ez with coordinate r. The frame ep is fixed to the pad body and shares its origin with es, its position is described by tilt angle and roll angle . Figure 3 shows the parameters used to describe the geometry of the pad. In the pad rigid body frame the position of a point on the pad surface n is described with constant coordinates x p y p z p 2 64 3 75 \u00bc Rpad sin p \u00f0 \u00de 1 2 qLpad Roff \u00fe Rpad 1 cos p \u00f0 \u00de\u00f0 \u00de 2 64 3 75 \u00f01\u00de Since the air film is thin (about 10 mm) the rotations of the pad are very small (order of magnitude 0.001 rad), and therefore the rotation matrix from e to ep can be written as epx epy epz 2 4 3 5 \u00bc 1 0 0 1 1 2 4 3 5 ex ey ez 2 4 3 5 \u00f02\u00de With this relation between the frame orientations, and the radial distance r between the origins of frames e and ep (see Figure 2), the surface point coordinates in the bearing centre inertial frame become x y z 2 64 3 75 \u00bc 1 0 0 1 1 2 64 3 75 x p y p z p 2 64 3 75 0 0 r 2 64 3 75 \u00f03\u00de And the velocity components are given by x _ y _ z _ 2 64 3 75 \u00bc 0 0 _ 0 0 _ _ _ 0 2 64 3 75 x p y p z p 2 64 3 75 0 0 _r 2 64 3 75 \u00f04\u00de An example model of the motion of the rigid rotor shaft is shown in Figure 4. If motion along the axial direction ey is restricted and the rotor has a circular cross section, the position of the rotor surface is fully determined by the four coordinates shown in the figure "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003319_cdc.2016.7799009-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003319_cdc.2016.7799009-Figure1-1.png",
        "caption": "Fig. 1. Logarithmic quantizer and a sector bound.",
        "texts": [
            " Moreover, for any s(0) \u2208 R(c), where R(c) := {s \u2208 R ns : V (s) \u2264 c} with c being a positive scalar such that R(c) \u2282 S, then s(k) \u2208 R(c), \u2200 k > 0, and limk\u2192\u221e s(k) = 0. Lemma 2 ([16]): Let P and G be real square matrices with P > 0 and G nonsingular. Then, the following holds: G\u2032PG \u2265 G+G\u2032 \u2212 P\u22121 . (11) Consider the closed-loop system of (1) with (3) and (5), namely : x+ = A(x)x +B(x)Q(Kx) . (12) It turns out that the logarithmic quantizer as defined in (6) is precisely bounded by a sector as illustrated in Fig. 1. Hence, the quantizer Q(v) can be alternatively defined as follows: Q(v) = v + q(v), q(v) : |q(v)| \u2264 \u03b4v , (13) with \u03b4 being as in (7). Then, the closed-loop system in (12) can be cast as follows: x+ = A\u0304(x)x +B(x)q(v) , (14) where A\u0304(x) = A(x) +B(x)K and q(v) is such that \u03b42v\u2032v \u2212 q(v)\u2032q(v) \u2265 0 . (15) Now, let the following Lyapunov function candidate: V (x) = x\u2032Px, P > 0 . (16) Then, for some scalar \u01eb3 > 0, we have that \u0393(x) = DV (x) + \u01eb3x \u2032x (17) = x\u2032 ( A\u0304(x)\u2032PA\u0304(x)\u2212P+\u01eb3I ) x+ 2q(v)\u2032B(x)\u2032PA\u0304(x)x + q(v)\u2032B(x)\u2032PB(x)q(v) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.20-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.20-1.png",
        "caption": "FIGURE 6.20",
        "texts": [
            " Space does not permit a description of the modelling of such systems here, but with ever more students becoming involved in motorsport this section will conclude with a description of the type of anti-roll bar model that might be included in a typical student race vehicle. A graphic for the system is shown in Figure 6.19. Graphic of anti-roll bar in typical student race vehicle. Provided courtesy of MSC Software. Modelling of anti-roll bar mechanism in student race car. The modelling of this system is illustrated in the schematic in Figure 6.20 where it can be seen that the anti-roll bar is installed vertically and is connected to the chassis by a revolute joint. The revolute joint allows the anti-roll bar to rock back and forward as the bell cranks rotate during parallel wheel travel but prevents rotation during opposite wheel travel when the body rolls. As the body rolls the torsional stiffness of the anti-roll bar, modelled with the rotational springdamper, resists the pushing motion of one push rod as the suspension moves in bump on one side and the pulling motion as the suspension moves in rebound on the other side"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001366_robio.2014.7090511-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001366_robio.2014.7090511-Figure3-1.png",
        "caption": "Fig. 3. Relationship between COG motion and ZMP position",
        "texts": [
            " According to linear inverted pendulum model (LIPM), to ensure robot\u2019s walking stability, acceleration of COG should be in accordance with planned data strictly in single support phase; But in double support phase, enlargement of support area and movement of ZMP between robot\u2019s feet provides possibility of realtime adjustment,namely the phase modification control method. As mentioned above, when ZMP locates in specific area of robot\u2019s foot, phase modification control can be taken and control parameters can be determined by ZMP\u2019s position and its relationship with COG. Figure.3 shows the relationship between COG motion and ZMP position. The biped system can be simplified as a inverted pendulum with an inertial flywheel and its mass is centered at the center of mass. Mlink is the torque due to flywheel rotation. fxand fz are forces the hinge exerts on the COG. PC(xC,yC,zC),PZ(xZ ,yZ ,zZ) are the positions of the COG and ZMP. l is the distance from the foot to COG. \u03b8c and \u03b8a are respectively,the flywheel and the leg angles with respect to vertical. The equations can be described as follows: Mlink + fx \u00b7 l \u00b7 cos\u03b8a = fz \u00b7 l \u00b7 sin\u03b8a (3) \u03b8\u0308c = 1 J Mlink (4) z\u0308C = fz m \u2212g (5) x\u0308C = ( fz(xC \u2212 xZ)\u2212Mlink) 1 m(zC \u2212 zZ) (6) where m and J are the mass and rotational inertia"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.31-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.31-1.png",
        "caption": "Fig. 2.31 Two-element truss structure with displacement boundary condition",
        "texts": [
            "213) Inverting the reduced stiffness matrix can be used to calculate the unknown displacements as: 68 2 Rods and Trusses \u23a1 \u23a3 u2X u2Y u3X \u23a4 \u23a6 = L E A \u23a1 \u23a2 \u23a3 E A L 0 0 \u2212 \u221a 3E A 5L 4 5 0 0 0 4 3 \u23a4 \u23a5 \u23a6 \u23a1 \u23a3 u 0 0 \u23a4 \u23a6 = u \u23a1 \u23a3 1 \u2212 \u221a 3 5 0 \u23a4 \u23a6 . (2.214) Reaction and rod forces can be obtained as described in part (a) as: R1X = \u22123 5 \u00d7 E Au L , R1Y = \u2212 \u221a 3 5 \u00d7 E Au L , R2X = 3 5 \u00d7 E Au L , (2.215) R3X = 0 , R3Y = \u221a 3 5 \u00d7 E Au L . (2.216) FI = 2 \u221a 3 5 \u00d7 E Au L , FII = \u221a 3 5 \u00d7 E Au L , FIII = 0. (2.217) 2.6 Example: Plane truss structure with two rod elements The following Fig. 2.31 shows a two-dimensional truss structure. The two rod elements have the same cross-sectional area A and Young\u2019s modulus E . The length of each element can be calculated based on the given dimensions in the figure. The structure is loaded by prescribed displacements u X and uY at node 2. Determine: \u2022 The global system of equations without consideration of the boundary conditions at node 1 and 3. \u2022 The reduced system of equations. 2.4 Assembly of Elements to Plane Truss Structures 69 \u2022 All nodal displacements"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000726_1.4031738-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000726_1.4031738-Figure2-1.png",
        "caption": "Fig. 2 NLR\u2019s Facility for Unmanned Rotorcraft Research project. Typical MR hub for a (small-scale) UAV helicopter (courtesy of NLR).",
        "texts": [
            " Now, for the case of a fully articulated rotor system, each rotor blade is attached to the rotor hub through a series of hinges, which allow each blade to move independently of the others. However, for small-scale helicopters, the rotor hub generally includes a pitch (feathering) hinge close to the shaft, and a lead\u2013lag hinge4 further outboard. Besides, the hub is typically not equipped with a flap hinge, the latter is often replaced by stiff rubber rings, hence a so-called hingeless flap mechanism, see Fig. 2. But for the purpose of helicopter flight dynamics modeling, it is a standard practice to model a hingeless rotor (and its flexible blades) as a rotor having rigid blades attached to a virtual hinge [49], the latter being offset from the MR axis. This virtual hinge is often modeled as a torsional spring, implying stiffness and damping.5 In order to simulate a generic flybarless small-scale helicopter MR, we have chosen to model it as an articulated P\u2013L\u2013F hinge arrangement. This chosen hinge configuration is particularly well suited for the case of small-scale helicopters. Indeed, it allows to keep the pitch and lag hinge offsets at their current physical values while replacing the rubber O-rings, see Fig. 2, by a virtual flap hinge (having stiffness and damping) outboard6 of the lag hinge. 3.1 Assumptions. The presented assumptions are valid for stability and control investigations of helicopters up to an advance ratio limit of about7 0.3 [51\u201353]. 3.1.1 Structural Simplifications (1) Rotor shaft forward and lateral tilt-angles are zero. Rotor precone is also zero. The blade has zero twist, constant chord, zero sweep, constant thickness ratio, and a uniform mass distribution. (2) We assume a rigid rotor blade in bending"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002862_s11082-016-0703-y-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002862_s11082-016-0703-y-Figure1-1.png",
        "caption": "Fig. 1 Schematic of the multicomponent-graded structure fabrication in Ti\u2013Al2O3 system by the LMD process of the ARTMC with: A micron and B nano sized alumina",
        "texts": [
            " The substrates were round plates with the 65 mm diameter and 5 mm height made of Ti-6Al-4 V. 447 Page 2 of 9 I. Shishkovsky et al. All the experiments were carried out using a HAAS 2006D (Nd:YAG, 4000 W, cw) with the laser beam delivery system, powder feeding system, coaxial nozzle, and numerically controlled 5-axes table. Some features of the equipment are reported in (Shishkovsky and Smurov 2012). The method of layerwise fabrication used in the present study was described thoroughly in (Shishkovsky et al. 2012) and is schematically presented in Fig. 1. Hatching distance was 2 mm, layer depth was *1 mm, and the powder feeding rate was *10 g/min. The layers were made out of Ti and Al2O3 powders on a substrate by the following strategy: the first layer was of titanium with 10 % Al2O3, the second one consisted of 80 % Ti ? 20 % Al2O3, the third layer\u2014of 70 % Ti ? 30 % Al2O3. In the first case (strategy A) we used the micron size alumina powder. In the second case (strategy B) it was the Ti ? nano Al2O3 mixture. Laser scanning speed was 500\u2013750 mm/min, laser power varied within range of 800\u20131000 W, and laser beam diameter was 3 mm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001393_ipec.2014.6869843-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001393_ipec.2014.6869843-Figure4-1.png",
        "caption": "Fig. 4. Overview of the rotor",
        "texts": [],
        "surrounding_texts": [
            "(PMSMs) with rare-earth permanent magnets (PMs) are most popular for automotive applications in which high torque density and high efficiency are required. Because the rare earth materials have risks of lacking stable supply, rare earth-free motors such as claw pole motors are widely studied. This paper presents the driving characteristics of a claw pole motor with no use of heavy rare-earth materials. In this motor the magnetic flux density in the air gap can be controlled arbitrarily by controlling field current. This paper shows that the variable field excitation can contribute to enlarge high efficiency range compared with fixed field excitation. Moreover, the electricity consumption driven in the Japan's fuel economy criteria (JC08 cycle) as EY and HEY applications are discussed.\nKeywords- Automotive electric motor, claw pole motor, field excitation control, rare-earth-free motor\nI. INTRODUCTION\nRecently the electrically-driven vehicles such as hybrid electric vehicles (HEVs) and electric vehicles (EVs) have become the key technologies of reducing greenhouse gas emissions[ I]. In these vehicles, the rare-earth permanent magnet motors, using Nd-Fe-B magnets including rare earth material such as neodymium (Nd), dysprosium (Oy) or terbium (Tb), are commonly used because of high magnetic density and high temperature capability. However, increasing demands of rare-earth material such as Dy and Tb will lead undersupply of these resources in the future. To solve this problem, rare-earth-free motors have been studied in automotive technologies.\nThere have been several rare-earth free motors which were investigated as conventional power trains [2]. As one example, induction motors (IMs) are commonly used for many application such as electric train. IMs are generally adopted longer axial length and distributed winding in order to increase torque capability. Thus, there have been studied for EV application by using IMs. Although, IMs seem not to be suitable for HEV application. Assuming mild HEV, motors are required to be sandwiched between the transmission and the engine. In this system, the requirements of the motors are shorter axial length and high torque density. By paying attention to this point, we have focused on claw pole motors (CPMs). CPMs can increase number of magnetic poles in the rotor easily\n978-1-4799-2705-0/14/$31.00 \u00a92014 IEEE\nbecause it is not necessary to substantially change the manufacturing. In addition, CPMs are able to field magnetomotive force arbitrarily by using control field current. As another type of the changing field magnetomotive force, there are also wound-field synchronous motors. Although, it has difficultness to increase rotor poles. Therefore, CPMs is expected to adopt HEV application.\nCPMs are able to control field magnetic force by means of controlling filed current arbitrary. When the field current is energized, the field flux is emerged. As the field magnetomotive force is increased, the more field current increases. In automotive application, the motor requires to keep high efficiency and high torque density. Although, the efficiency map of conventional PMs are fixed field magnetomotive force. In such as PMs, it is necessary to control magnetic flux weakening in high rotation to reduce high line voltage. Thus, copper loss is increased and iron loss is kept highly. On the other hand, claw pole motors are easily to reduce magnetic flux not to use flux weakening control. It is possible to achieve the effect of magnetic flux weakening by decreasing field current. Therefore, claw pole motors are able to keep high efficiency in any operational points.\nOur final target was achieved that the torque density was more than SNm/kg, motor efficiency was more than 90% in order to be close to PMSM performance. We achieved at the first stage that torque density was 3Nm/kg, motor efficiency was 80% which was based on alternators in the 1st prototype (CPM1). However, CPMI was required to reduce weight and size to be close to PMSM performance at the same torque. Based on the results of CPM 1, the 2nd prototype (CPM2) was made in order to improve efficiency and torque density. Firstly, with the goal of high efficiency, the surfaces of the claws were changed from solid core to laminated core for the purpose of decreasing eddy current loss in the rotor. Secondary, targeting of high output density, each claw was inserted ferrite magnets to decrease magnetic saturation. As a result, we achieved that torque density was 4Nm/kg, motor efficiency was 89% [3]. As a way of more improvements, we made the 3rd prototype (CPM3) so that the rotor diameter of CPM3 was larger than that of CPM2 to decrease magnetic saturation between these claws. As a result, we achieved the final targets, SNm/kg of torque density and 90% of motor efficiency which are nearly equal to the performance of PM motors[4].\n1892",
            "In this paper, it presents that the validity of the variable field excitation control of CPM3 which can be controlled field magnetomotive force arbitrary. The difference of the motor efficiency between variable field excitation and fixed field excitation is shown using FEA analysis. Moreover, it is shown that the simulation of the EV and HEY application in the case of lC08 cycle.\nFigure I shows magnetic circuit of CPM2 and CPM3. As shown in Fig.l, there are two air gaps in this motor. One of them is between stator and rotor. Another is between inside the claw and field unit. When the field coil is energized, it supplies the field coil without slip rings throughout the stator bracket. The field flux passes between field coil and inside the rotor. Moreover, the ferrite magnets are located between each claw. Here, it is noted that the magnetic polarity of the ferrite magnets was opposite to the field flux. The opposite polarity can suppress the flux leakage in the rotor so that magnet field flux cancels flux leakage in the rotor. Therefore, it is expected to increase maximum torque while it suppress magnetic saturation.\nTable I shows specifications of proposed CPM2 and CPM3. As a way of increasing maximum torque, the number of rotor poles was increased from 24 to 32 so that CPM3 could decrease magnetic saturation in the rotor. Thus, the magnet weight of CPM3 was increased 1.3 times than that of CPM2. The reason for this is that the increased number of ferrite magnets could suppress flux leakage effectively by increased pole number.\nMoreover, the diameter of the rotor was increased so that CPM3 could increased air gap flux density. The stack length of the rotor was longer than that of the stator so that the axial flux could pass through the rotor. In addition, the concentrated winding was aimed to decrease copper loss compared to the distributed winding. Moreover, these claws were adopted laminated core so that the core could decrease eddy current loss. Most eddy current loss was generated on the surface of the claws. The other parts that faced stator core was composed of solid core.\nFig. 2. Cross-section of CPM3\nField coil Field bracket\nFig. 3. Overview of the field unit",
            "The field unit was inserted inside the rotor. It is noted that the field unit was attached with the stator bracket. Thus, field unit was supplied field magnetomotive force while it was fixed with stator bracket.\nThe specification shown above, CPM3 has an advantage that field loss can be decreased arbitrary by controlling field current optimally.\nIII. ANALYSIS CONDITION OF THE PROPOSED CPM\nWe evaluated two conditions in order to compare these loss in FEA. Condition I and II are assumed for assist of acceleration (I OOOmin-l, 65Nm) and steady state (3000 min-I, 25Nm). In these conditions, the effect of the efficiency were compared variable field excitation with fixed field excitation. The assist of acceleration required high magnetomotive force so that the motor can output required high torque. The steady state requires low field excitation so that the field excitation decrease iron loss in the high rotational speed.\nAs previous described, field excitation is necessary to control arbitrary in each operational points. Here, it is shown that the motor efficiency map are compared variable field excitation with fixed field exaction.\nUsing motor efficiency map in FEA results, electricity consumption was simulated in the case of JC08 cycle.\nTABLE II shows the simulated condition for EV application. In this application, the electricity consumption was calculated by all operational points in lC08 cycle. In this TABLE, car weight was considered 1300kg included buttery weight. The gear ratio was set at 6. Besides, the gear efficiency was considered l.0. The conditions of two field excitation were considered variable and fixed conditions. In the simulation, the maximum value of the fixed field excitation was set 4500AT so that the motor could output maximum torque in JC08 cycle. For example, when the gear ratio is 6, maximum torque is 90Nm.\nUsing motor efficiency map in FEA results, electricity consumption was simulated when the maximum output was 10kW assistance. TABLE III shows the simulated condition of 10kW assistance for HEV. The simulation was taken four conditions into consideration. These conditions compared the loss dependence of the gear control and field excitation. Two conditions in terms of the mechanical gear were compared. The variable gear was considered from 2 to 10. The gear could change so as to minimize the motor loss. The fixed gear ratio was set at 6. In this simulation, the gear efficiency was 1.0. In the field condition, the conditions of field excitation were variable and fixed. The variable excitation was set optimally within 4500A T. It is for the reason that the maximum value could output maximum torque.\nFigure 5 shows comparison of total loss with field excitation control and fixed field excitation. As for condition I, the total loss was decreased 230W with variable field excitation control. The field loss was decreased 360W, nevertheless, the copper loss and iron loss were increased 80W and 50W. It is considered that the armature current was increased so as to output the same torque. Thus, the armature current was increased. Field flux was also decreased. In the iron loss, it could be seen that the iron loss was increased though the fixed field excitation control was decreased. It suggests that alternating magnetic field was increased because the current amplitude was increased.\nIn condition II, condition II could be found that the total loss was decreased 540W by variable field excitation control. Compared to fixed field excitation control, copper loss and iron loss were increased as well as condition I."
        ]
    },
    {
        "image_filename": "designv11_64_0000469_978-3-319-21118-3_19-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000469_978-3-319-21118-3_19-Figure13-1.png",
        "caption": "Fig. 13 Rotors thrusts in forward flight",
        "texts": [
            " The detailed analysis indicated that the disturbances resulted from non uniform and asymmetric variations of thrust of rotors. In the next tests the flight level and azimuth controllers were improved. An additional attitude feedback loop was applied to compensate fluctuations in altitude, which were mainly caused by the rotors inclination angle, Fig. 12. The rotors angular velocity increment \u0394\u03c9 was then calculated using actual rotor attitude gained by factor kc to fit with actual controller settings. The constant altitude in a forward flight is possible if a rotor thrust TF at specific inclination angle \u03b3 (Fig. 13) fulfills the condition: TF cos c\u00f0 \u00de \u00bc TH ; \u00f09\u00de where TH is the rotor thrust at hovering conditions. The inclination angle can be estimated as: cos c \u00bc cosU cosH. \u00f010\u00de If a rotor thrust is expressed as a square of function of rotor angular rate: Ti \u00bc CTX 2 i ; \u00f011\u00de where CT is thrust constant, and the relationship between rotor angular rates at hover and forward flight conditions are: XF \u00bc XH \u00fe DX; \u00f012\u00de then the Eq. (9) can be rewritten as: XH \u00fe DX\u00f0 \u00de cos c\u00f0 \u00de \u00bc XH: \u00f013\u00de The increment of rotor angular rates \u0394\u03a9 for forward flight calculated as: DX \u00bc ffiffiffiffiffiffiffiffiffi cos c p cos c 1 XH: \u00f014\u00de The azimuth controller was enhanced by additional attitude feedback loop (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.34-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.34-1.png",
        "caption": "Fig. 3.34 Portal frame structure",
        "texts": [
            "2 99 ) 162 3 Euler\u2013Bernoulli Beams and Frames The size of some structures can be reduced by consideration of symmetry and antisymmetry boundary conditions as shown in Fig. 3.33. The case of symmetry requires that the geometry and the load case is symmetric with respect to a certain plane while the case of anti-symmetry requires a symmetric geometry and anti-symmetric loading and results with respect to the same plane. A systematic summary of symmetric and anti-symmetric boundary conditions is given in Table3.16. 3.4 Assembly of Elements to Plane Frame Structures 163 The portal frame structure shown in Fig. 3.34 is loaded by a constant distributed load q and a single horizontal force F. The three parts of the frame have the same length L, the same Young\u2019s modulus E, the same cross-sectional area A, and the same second moment of area I . Determine (a) \u2022 the elemental stiffness matrices in the global X\u2013Z system, \u2022 the reduced system of equations, \u2022 all nodal displacements and rotations, \u2022 all reaction forces. (b) Consider now the case F = 0, i.e. only the distributed load is acting on the frame. Develop a simpler model under consideration of the symmetry and determine the quantities as requested in part (a)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.37-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.37-1.png",
        "caption": "FIGURE 6.37",
        "texts": [
            " The opposite arrangement e a higher value on centre and a lower level towards the limits could be argued as desirable to increase agility at speed and reduce the burden on power steering systems when parking. The authors suspect the advantages and disadvantages are probably larger in the minds of the steering system vendors than of the customers. Using the MBS approach the steering ratio can be investigated through a separate study carried out using the front suspension system connected to the ground part instead of the vehicle body. The modelling of these two subsystems, with only the suspension on the right side shown, is illustrated in Figure 6.37. The approach of using a direct ratio to couple the rotation between the steering column and the steer angle of the road wheels is common practice in simpler Front suspension steering ratio test model. models but may have other limitations in addition to the treatment of the ratio as linear: 1. In the real vehicle and the linkage model the ratio between the column rotation and the steer angle at the road wheels would vary as the vehicle rolls and the road wheels move in bump and rebound. 2. For either wheel the ratio of toe out or toe in as a ratio of left or right handwheel rotation would not be exactly symmetric",
            " direction of handwheel rotation if the behaviour is to be modelled. It should also be noted that compliance in the steering rack or rotational compliance in the steering column could be incorporated if it adds value to the analysis. In the followingexample thegeometric ratio between the rotation of the steeringcolumn and the travel of the rack is already known, so it is possible to apply amotion input at the rack to ground joint that is equivalent to handwheel rotations either side of the straight ahead position. The jack part shown in Figure 6.37 can be used to set the suspension height during a steering test simulation. Typical output is shown in Figure 6.38 where the steering wheel angle is plotted on the x-axis and the road wheel angle is plotted on the y-axis. The three lines plotted represent the steering ratio test for the suspension in the static (initial model set up here), bump and rebound positions. Having decided on the suspension modelling strategy and how to manage the relationship between the handwheel rotation and steer change at the road wheels, the steering inputs from the driver and the manoeuvre to be performed need to be considered"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002311_imece2016-65745-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002311_imece2016-65745-Figure5-1.png",
        "caption": "FIGURE 5: TORSIONAL BENDING LOAD STEPS FOR P3G CONNECTION",
        "texts": [
            " The body was meshed with lower order hexahedron mesh (Solid 185) with contact elements of 0.35 mm. The total number of elements varied according to the fit, averaged approximately 210,000. The meshing for P3G and P4C shaft hub connections are shown in Fig. 3 and Fig. 4 respectively. iii) Torsional bending load The analysis was performed by applying torsional bending load to the polygonal shaft hub connection. Only half of the connection was modeled due to symmetry along the axis with hub at the center of the connection as shown in Fig. 5. A torque of 80 Nm and a corresponding bending load of 4600 N was considered for the analysis to emulate the load from a spur gear. Initially, a rotating bending load was applied at 10\u00b0 difference for 120\u00b0 for P3G and 90\u00b0 for P4C connection along with static torsion to find the critical loading position as shown in Figs. 5 and 6 since the shaft is not radially symmetrical. 3 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90996/ on 07/24/2017 Terms of Use: http://www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001595_amm.590.394-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001595_amm.590.394-Figure1-1.png",
        "caption": "Fig. 1. Seven DOFs Dynamic model of the vehicle.",
        "texts": [
            " By making full use of the excellent control capability of electric drive systems, EVs can not only be pollution-free, but also be able to achieve better maneuverability that cannot be accomplished by the traditional ICEVs. In this paper, the stability of four-wheel independent driving EVs is studied and the primary factors under the low-adhesion, sharp-corner road conditions are investigated and evaluated. The results in this research provide a solid guidance to achieve better maneuverability for the four-wheel independent driving electric vehicles, which will definitely accelerate the process of massive production in the near future. A seven degree of freedom dynamic model is illustrated in Fig. 1, in which seven DOFs are considered, including the longitudinal motion, the lateral motion, the yaw motion and the rotation of the four wheels. xrrxrlxfrxflx FFFFvumma +++=\u2212= '')( \u03b3 (1) yrryrlyfryfly FFFFuvmma +++=+= '')( \u03b3 (2) )( 2 )( 2 )()( '''' xrrxrl r xfrxfl f yrryrlryfryflfz FF D FF D FFlFFlI ++\u2212++\u2212+=\u03b3 (3) All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003236_detc2016-59619-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003236_detc2016-59619-Figure2-1.png",
        "caption": "FIGURE 2. A SIMPLE 5 DOFS CHAIN WITH ONE BRANCH",
        "texts": [
            "org/about-asme/terms-of-use where x\u03040 is the initial COM of the serial chain, and \u03be\u0302 \u2032 i is the twist for the ith joint of the new equivalent chain, \u03be\u0302 \u2032 i can be computed using the following equations: e\u03b8n\u2212 j \u03be\u0302 \u2032 n\u2212 j = Mn\u22121Mn\u22122 \u00b7 \u00b7 \u00b7Mn\u2212 je\u03b8n\u2212 j \u03be\u0302n\u2212 j M\u22121 n\u2212 j \u00b7 \u00b7 \u00b7M \u22121 n\u22122M\u22121 n\u22121 (2) M j = I3\u00d73 m j+1 \u2211 j i=1 mi c\u0304 j+1 0 \u2211 j+1 i=1 mi \u2211 j i=1 mi  \u03c1 j = \u2211 j i=1 mi \u2211 j+1 i=1 mi . The similarity is readily observed between (1) and the equation for the forward kinematics of a n link serial chain. As a consequence, we formulate the method to construct the SESC model for a serial kinematic chain using POE formula, which is intuitive and time saving. Then, we study the SESC model for the branched chain, which is depicted in Figure 2. Similarly, we first consider the COM of only one link L5, which can be represented as [ x\u03045 1 ] = e\u03b85\u03be\u03025 [ c\u03045 1 ] , Then to make the branched chain work as a serial chain, we virtually link joint 5 with L4. The new COM equation then becomes [ x\u03045 1 ] = e\u2212\u03b84\u03be\u03024e\u2212\u03b83\u03be\u03023 e\u03b85\u03be\u03025 [ c\u03045 1 ] Finally, by following the computing direction along the chain shown as the dashed arrow in Figure 2, we get the final expression of COM for this branched chain, [ x\u0304 1 ] = e\u03b81\u03be\u03021e\u03b82\u03be\u0302 \u2032 2 e\u03b83\u03be\u0302 \u2032 3 e\u03b84\u03be\u0302 \u2032 4 e\u03b85\u03be\u0302 \u2032 5 [ x\u03040 1 ] where e\u03b85\u03be\u0302 \u2032 5 = M1M2 \u00b7 \u00b7 \u00b7M4e\u03b85\u03be\u03025\u2212\u03b83\u03be\u03023\u2212\u03b84\u03be\u03024M\u22121 4 \u00b7 \u00b7 \u00b7M \u22121 2 M\u22121 1 e\u03b85\u2212 j \u03be\u0302 \u2032 5\u2212 j =M5\u22121M5\u22122 \u00b7 \u00b7 \u00b7M5\u2212 je\u03b85\u2212 j \u03be\u03025\u2212 j M\u22121 5\u2212 j \u00b7 \u00b7 \u00b7M \u22121 5\u22122M\u22121 5\u22121, j = 1,2,3. This simple example is effective, because it may be formulated to directly produce the SESC model for the simple branched chain, primarily by following the computing direction. Kinematic analysis of the hexapod robot manipulation Hexapod robots have the highest efficiency for statically stable walking and can use one, two or three legs to function as hands [13,14]",
            " We can rewrite Equation (3) as, Vpo = Jopt [\u03b8\u03071, \u00b7 \u00b7 \u00b7 , \u03b8\u03076] T (4) Jopt is defined as the operating Jacobian matrix, which shows the kinematic relation between the object velocity and the hexapod robot joint angles limited by grasping constraints. We can consider the manipulation model as a 25 DoFs chain with six branches by regarding the body of the robot as six equal mass links connected by the same joint, and by treating the grasped object as being connected to the tip of leg1 with two orthogonal joints. Then following the similar procedure depicted in Figure 2, we construct the SESC model for the manipulation. However, here we model each branch by using the SESC method, then combine the virtual COM of each branch to get the COM of the system by a classic method. To implement the real time control, a classic velocity based control method is adopted and the Jacobian matrix for the COM of system is provided. The position of the COM could be expressed as, x\u0304 = 1 mr +mo +\u2211 6 i=1 mi (mr x\u0304r +mox\u0304o + 6 \u2211 i=1 mix\u0304i) where, x\u0304i is the location of COM for leg i as shown in Figure 3, and it is modeled by the end-point position of a 3 DoFs of the SESC model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000464_b978-0-08-096532-1.00508-2-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000464_b978-0-08-096532-1.00508-2-Figure15-1.png",
        "caption": "Figure 15 Ni-based superalloy turbine blades solidified as (a) equiaxed grains, (b) columnar grains, and (c) a single crystal and (d) schematic illustration of the as-cast single-crystal structure. After Reed, R. C. The Superalloys: Fundamentals and Applications; Cambridge University Press, 2006.",
        "texts": [
            "2 Grain Selection during Single-Crystal Casting The development of single-crystal technology for the fabrication of Ni-based superalloy turbine blades with complex geometries has been a critical factor in the progressive development of gas turbine power and efficiency. Compared with conventional casting and directional solidification casting, the benefits of single-crystal structure include: 1. Elimination of grain boundaries that exist in a polycrystalline morphology, such as equiaxed or columnar, resulting in greatly enhanced creep rupture and fatigue properties (Figure 15) (11,26,57\u201362). 2. Removal of grain boundary strengthening additions, such as B and Hf, which raises the incipient melting temperature and improves the operating window in the solution heat treatment (26,63). 3. Improvement of creep ductility and thermal fatigue resistance by orientating the preferred <001> crystallographic growth direction, which coincides with the elastically soft direction, parallel to the maximum loading direction (64\u201366). In order to grow single-crystal component, a grain selector is implemented into the directional solidification process in a modified Bridgman furnace to ensure that only one grain can survive in the final structure (Figure 16) (25,26)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000968_20140824-6-za-1003.01643-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000968_20140824-6-za-1003.01643-Figure1-1.png",
        "caption": "Fig. 1. Schematic view of the UAV path toward the objective circle",
        "texts": [
            " Finally, the UAV leaves this straight line and makes a loitering to converge to the objective circle. This methodology is discussed in detail in Subsection 3.1. The problem is extended to multi-UAV scenario in Subsection 3.2. The worst case for collision avoidance scheme is that all the UAVs are at the same height. Therefore, we will consider the motion of the UAVs in the yx \u2212 plane. As mentioned previously, in our proposed methodology, the path to bring a UAV to the objective circle consists of 3 segments, shown in Fig. 1. The green square in Fig. 1 will be later defined and employed in a manner such that initial deadlocks can be avoided. Also, arr\u03b1 will be later defined. An important definition is given here: Definition 1: Loitering circle of the UAV at a given point is the loci of all points that would be occupied by the UAV if it started loitering with its maximum RoT. The loitering circle corresponding to a clockwise loitering is called the right loitering circle and the one corresponding to a counterclockwise loitering is called the left loitering circle",
            " The following actions are executed in each state: Loitering state: maxmin vvvv c \u2264\u2264= and max\u03c9\u03c9 \u2212= Cruise state: cvv = and \u03c9 = 0 Final Approach state: cvv = and max\u03c9\u03c9 \u2212= On the Objective Circle state: cvv = and max\u03c9\u03c9 \u2264= o c R v The transitions can be viewed in the state diagram presented in Fig. 2. In the first state, i.e. the Initial Loitering state, the UAV starts loitering with its maximum RoT ( max\u03c9\u2212 ) in the clockwise direction, with constant forward velocity cv ( maxmin vvv c \u2264\u2264 ). The tangent line connecting the center of the objective circle to the initial loitering circle is the second part of the path toward the objective circle, i.e. the Cruise state. The point where the initial loitering circle is patched to the cruise path is called the Leave Point (see Fig. 1). In the Cruise state, forward velocity input is again cv and 0=c\u03c9 . At the end of the cruise state, the UAV reaches a point called Final Cruise Point (FCP), at which the loitering circle of the UAV is tangent to the objective circle. At this point, the UAV starts the final part of the path, i.e. the Final Approach state. In this state, the forward velocity command remains unchanged ( cv ) and the angular velocity is max\u03c9\u2212 . The Final Approach state ends when the UAV reaches the objective circle and then it starts circulating it",
            " Also, in the transitions, the expression \u201cflight corridor is blocked\u201d means that the UAV\u2019s flight corridor is already blocked by the safety circle of another UAV or it is in one of the three defined scenarios of potential collision and it has lower priority in relation to the other UAVs in its detection range. Convergence to the objective circle, in multi-UAV scenario is discussed in the following paragraphs. First, an assumption is made such that the UAVs are not \u201cborn\u201d in an initial deadlock configuration. Assumption 1: Consider a square with the edge length ( )UAVl rR +2 , with its center coincident with that of the initial loitering circle of the UAV and one of its edges parallel to the UAV\u2019s flight corridor (See Fig. 1). It is clear that, in the Initial Loitering state, the UAV is confined to the space encompassed by the defined square. It is assumed that the different squares corresponding to different UAVs do not initially overlap each other. Assumption 1 is necessary to ensure that the UAVs do not inevitably collide with each other, due to inappropriate initial conditions. Proposition 2: Consider maxNN \u2264 nonholonomic fixed-wing UAVs, initially out of the objective circle, with detection range ( ) UAVl rR *2224 ++ , with the initial conditions not violating Assumption 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003817_tia.2014.2322145-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003817_tia.2014.2322145-Figure6-1.png",
        "caption": "Fig. 6. Phasor diagram for rotor position estimation scheme.",
        "texts": [
            " The actual active and reactive powers are calculated as, 2 ( ) 3 P v i v is s s s s\u03b1 \u03b1 \u03b2 \u03b2= \u2217 + \u2217 (15) 2 ( ) 3 Q v i v is s ss s\u03b1 \u03b1\u03b2 \u03b2= \u2217 \u2212 \u2217 (16) where, vs\u03b1, vs\u03b2 are two phase stationary frame voltages calculated as, 1 1 3 3 v v vscs sb\u03b1 = \u2212 + (17) v vsas\u03b2 = (18) These three phase stator phase voltages (vsa, vsb, vsc) are estimated from two sensed line voltages (vab, vbc) as, 2 1 1 1 1 3 1 2 vsa v ab v sb v bc vsc           = \u2212         \u2212 \u2212    (19) where is\u03b1, is\u03b2 are the two phase stationary frame currents are calculated from the sensed stator currents (isa, isb) by using Clarke\u2019s transformation as, 1 1 3 3 i i is cs s b\u03b1 = \u2212 + (20) i isas\u03b2 = (21) Stator flux in rotor reference frame can be calculated as, { }cos( ) sin( )( ) r jm est m estjs s s \u03b8 \u03b8\u03c8 \u03c8 \u03c8\u03b1 \u03b2 \u2212= + (22) The stator flux linkages (\u03a8s\u03b1, \u03a8s\u03b2) are calculated as, ( )v R i dtss s s\u03c8 \u03b1 \u03b1 \u03b1= \u2212\u222b (23) ( )v R i dtss s s \u03c8 \u03b2 \u03b2 \u03b2 = \u2212\u222b (24) (\u03b8m)est is the estimated angle between stator and rotor. Unit templates of slip angle (cos (\u03b8m)est, sin (\u03b8m)est) are found from rotor position estimation algorithm given in the following section. In this algorithm, rotor current (ir) makes an angle \u03b8s from the stator co-ordinate system. Fig. 6 shows the same rotor current (ir) makes an angle \u03b8r from the rotor co-ordinate system. The angle between stator and rotor are calculated as (\u03b8m)est = (\u03b8s-\u03b8r). The schematic diagram of the sensorless scheme is shown in Fig. 7. Rotor currents (ira, irb) are sensed and transformed into two phase system using Clarke\u2019s transformation as, 1 1- 3 3 i i ircr rb\u03b1 = + (25) i irar\u03b2 = (26) Rotor current magnitude and unit templates of rotor currents aligned to rotor axis are calculated as, ( ) 1 22 2 i i ir r r\u03b1 \u03b2 = + (27) cos ir r ir \u03b1\u03b8 = (28) sin i r r ir \u03b2 \u03b8 = (29) The rotor current unit templates (cos \u03b8s, sin \u03b8s ) aligned to 0093-9994 (c) 2013 IEEE"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000469_978-3-319-21118-3_19-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000469_978-3-319-21118-3_19-Figure4-1.png",
        "caption": "Fig. 4 Propeller velocity vector in body reference frame",
        "texts": [
            " Axial inflow positive velocities W appear for instance when a quadrotor is climbing vertically. In results presented, rotor thrust is very sensitive to variation of inflow velocity, so keeping thrust constant for all inflow velocities results in a significant overestimation, even for axial flight. It can be also noticed that the rotor torque is not so much sensitive and axial inflow velocity variations do not change it very much. Not only quadrotor linear velocities contribute to change of inflow velocities. Propellers are placed not in the quadrotor center of gravity (Fig. 4) so also body angular rates may contribute to inflow velocities: vti \u00bc vb \u00fe xb rti \u00f04\u00de where vb; xb\u2014vectors of quadrotor velocities and rates in the body reference frame, rti\u2014position vector of a propeller in the body reference frame (Fig. 4). Neglecting this effect for the quadrotor, which is a very agile vehicle, limits simulation model to hover and low speed flights. As another effect resulting from (4) in a maneuver flight, each rotor is subjected to different inflow and thus generates different loads. These observations were taken into account in the WUT developed model. The quadrotor propellers loads consist of a full set of six components\u2014three forces and three moments of forces, calculated using Blade Element Method. First, the velocities at the blade cross-section are expressed as functions of body linear velocities and angular rates, including also velocities resulting from rotors angular velocities and induced velocity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.3-1.png",
        "caption": "FIGURE 6.3",
        "texts": [
            " Although this would not add DOF, the model would be less efficient through the introduction of additional equations representing the extra body and the fix joint constraint. The use of fix joint constraints may also A detailed multi-body systems vehicle model. Provided courtesy of MSC Software. Vehicle body reference frame. introduce high reaction moments that would not exist in the model when using elastic mounts distributed about the mass. An example of a vehicle body referenced frame O2 located at the mass centre G2 for Body 2 is shown in Figure 6.3. For this model the XZ plane is located on the centre line of the vehicle with gravity acting parallel to the negative Z2 direction. Using an approach where the body is a single lumped mass representing the summation of the major components, the mass centre position can be found by taking first moments of mass, and the mass moments of inertia can be obtained using the methods described in Chapter 2. From inspection of Figure 6.3 it can be seen that a value would exist for the Ixz cross product of inertia but that Ixy and Iyz should approximate to zero given the symmetry of the vehicle. In reality there may be some asymmetry that results in a CAD system outputting small values for the Ixy and Iyz cross products of inertia. The dynamics of the actual vehicle are greatly influenced by the yaw moment of inertia, Izz, of the complete vehicle, to which the body and associated masses will make the dominant contribution. A parameter often discussed is the ratio k2/ab, sometimes referred to as the \u2018dynamic index\u2019, where k is the radius of gyration associated with Izz and a and b locate the vehicle mass centre longitudinally relative to the front and rear axles respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001972_ilt-03-2015-0034-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001972_ilt-03-2015-0034-Figure6-1.png",
        "caption": "Figure 6 Computation domain",
        "texts": [
            " Turbulent flow behavior in plain journal bearing Nadia Bendaoud, Mehala Kadda and Abdelkader Youcefi Industrial Lubrication and Tribology Volume 68 \u00b7 Number 1 \u00b7 2016 \u00b7 76\u201385 D ow nl oa de d by F lo ri da A tla nt ic U ni ve rs ity A t 2 2: 59 0 6 M ar ch 2 01 6 (P T ) The physical domain shown in Figure 3 is discretized into volume tetrahedral elements form. Figure 5 corresponds to the mesh of the hydrodynamic bearing. This mesh has 20 nodes along the bearing and 103 nodes for its circumference. The diameter of the supply ports is discretized into 14 nodes. The length of the groove consists of 72 nodes, 16 nodes for its width and 5 nodes along the depth of the groove. The film thickness is discritized by 25 nodes. The mesh consists of elements 30,433 and nodes 74,163. For the boundary conditions (Figure 6), a supply pressure is applied in the three supply ports while assuming that the inlet and outlet of the fluid with a temperature of 40\u00b0C, and in the pressure of 0.08 MPa. Except the shaft is considered a rotating element. However, the bushing is recessed on its outer diameter. The loading system is simulated by an orifice of diameter 14 mm, the direction of the fluid inlet is the same radial load direction with variable pressures calculated by the ratio of the load on the bearing surface, with temperature of 40\u00b0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001605_naps.2015.7335151-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001605_naps.2015.7335151-Figure3-1.png",
        "caption": "Fig. 3. Proposed Single Equivalent Circuit of a Salient-Pole Synchronous Machine",
        "texts": [
            " However, the reluctance power term is not readily identifiable in this classical phasor diagram. In the next section, we propose a new singlecircuit representation of a salient-pole synchronous machine along with a modified phasor diagram where the two terms of the internally-developed electromagnetic power are readily identifiable. V. REVISED PHASOR DIAGRAM By defining the following reactances, + =  + 2 (47) \u2212 =  \u2212 2 (48) the steady-state voltage equations (41)-(42) can be manipulated to yield the following phasor equation which is represented by the single equivalent circuit shown in Figure 3: \u2212\u2192  = ( + +) \u2212\u2192  + \u2212 \u2212\u2192 \u2217 +  (49) If we define an internal voltage \u2212\u2192  as the sum of the de- pendent voltage source (\u2212 \u2212\u2192 \u2217) and the independent voltage source (), \u2212\u2192  =  + \u2212 \u2212\u2192 \u2217 =  + \u2212( \u2212 ) =  + \u2212 +\u2212 (50) then we can easily verify that the internally-developed electromagnetic power  is equal to the real power absorbed by this internal voltage by expanding the following equation  = <{\u2212\u2192  \u2212\u2192 \u2217} = <{\u00a1 + \u2212 +\u2212 \u00a2 ( \u2212 )} = 2\u2212 + (51) = ( \u2212) + (52) A phasor diagram for the single equivalent circuit can be drawn as shown in Figure 4 where the stator resistance has been omitted for simplicity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001244_esda2014-20131-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001244_esda2014-20131-Figure1-1.png",
        "caption": "Fig. 1 Configuration of the externally pressurized air bearing",
        "texts": [
            " Pressure distribution which is important parameters of externally pressurized gas lubricated bearing was modeled and solved using numerical solution method. The flow between bearing and rotor was modeled using Reynolds's Equation which is known fluid equation of motion. And this model was solved using by Alternating Direction Implicit (ADI) numerical solution method. Dynamics of rotor was investigated for different high (increased) clearances and different supply pressure. 2. MATHEMATICAL MODEL The bearing-rotor system, orifices on the bearing and coordinate system used in the modelling are shown in Fig. 1. Radial motion of the rotor which is supported by externally pressurized gas bearing could be modeled in two degree of freedom using cartesian coordinates and given in equation 1 and 2. In this model it is assumed that rotor does only have radial motion. Right hand side of the following equations is zero because the rotor was assumed balanced and there is no external force applied it. 2 L/R 2 a 0 0 mx p R P( , ) cos d d 0          (1) 2 L/ R 2 a 0 0 my p R P( , )sin d d 0          (2) The pressure distribution between rotor and bearing is modeled by using Reynold\u2019s equation; 3 3 0 0h p p h p p 12R T m 12 (ph) x x z z t 6U (ph) x                             (3) and here it is assumed that; (1) The flow between rotor and bearing is isothermal"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.40-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.40-1.png",
        "caption": "Fig. 3.40 Plane frame structure composed of generalized beam elements",
        "texts": [
            " \u2022 Explain in words the difference between an Euler\u2013Bernoulli beam element and a generalized beam element in regards to the nodal unknowns. \u2022 State the DOF per node for a generalized beam element in a plane (2D) problem. \u2022 State the DOF per node for a generalized beam element in a 3D problem. \u2022 Figure3.39a shows schematically a cantilevered Euler\u2013Bernoulli beam. In a finite element approach, such a beam can be modeled based on one-dimensional beamelements (Fig. 3.39b), two-dimensional plane elasticity elements (Fig. 3.39c), or three-dimensional solid elements (Fig. 3.39d). State for each approach one advantage. \u2022 Figure3.40 shows a plane frame structure which should be modeled with three generalized beam (I, II, III) elements. State the size of the stiffness matrix of the non-reduced system of equations, i.e. 172 3 Euler\u2013Bernoulli Beams and Frames without consideration of the boundary conditions. What is the size of the stiffness matrix of the reduced system of equations, i.e. under consideration of the boundary conditions? 3.11 Cantilevered beam with a distributed load: analytical solution Calculate the analytical solution for the deflection uz(x) and rotation \u03d5y(x) of the cantilevered beam shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000913_j.mechmachtheory.2014.07.010-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000913_j.mechmachtheory.2014.07.010-Figure2-1.png",
        "caption": "Fig. 2. Transmission dynamic model.",
        "texts": [
            " Obtained analytical expressions and diagrams allow an optimal choice of the operating parameters. Thus the normal operation of the band saw machines can be guaranteed. icho_marinoff@abv.bg. 0 2. Expose 2.1. Principle scheme of the band saw machine The scheme of the band sawmachine is shown in Fig. 1 [1\u20136]. The following symbols are defined: 1, 2, 5, 6\u2014 belt pulleys, E\u2014 electric motor, 3 and 4 \u2014 leading wheels, A \u2014 band-saw blade, and 7 and 8 \u2014 chain-wheels. 2.2. Transmission dynamic model The dynamic model is shown in Fig. 2 [7,8]. This model is used to solve the problems. We choose the following coordinate systems [7,8]: Fixed coordinate system O3xyz, moving coordinate system O3x1y1z1, which moves along with the driving wheel. In the initial moment (\u03c6 = 0) the axes of the two coordinate systems coincide. Coordinate system C3x\u2032y\u2032z\u2032, beginning at the centre of mass C3. Its axes are parallel to the axes of the moving coordinate system. We use a fourth coordinate system C3\u03be\u03b7\u03c2. The axes of this coordinate system are principal axes of inertia of the disk"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.57-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.57-1.png",
        "caption": "Fig. 3.57 Beam structure with a gap in the middle and vertical point load at the right end",
        "texts": [
            " Model the beam with two elements to determine: 180 3 Euler\u2013Bernoulli Beams and Frames \u2022 the unknown rotations and displacements at the nodes, \u2022 the reaction forces at the supports, \u2022 the vertical deflection in the middle of the section with the distributed load, i.e. X = 1 2LI, \u2022 the bending moment and shear force at the midpoint of the section with the distributed load, i.e. X = 1 2LI. \u2022 Improve the approximate solution for uZ(X = 1 2LI) by subdividing the section with the distributed load in two elements of equal length. 3.32 Beam structure with a gap The beam shown in Fig. 3.57 is loaded by a single vertical force F at its right-hand end. The bending stiffness EI is constant and the total length of the beam is equal to 2L. In the middle of the entire structure, there is a pap of length \u03b4 between the beam and a simple support. Model the beam with two elements to determine: \u2022 the force to close the gap, \u2022 the deflection and rotation at the free end, i.e. X = 2L, as a function of the increasing force F, \u2022 the maximum normalized stress \u03c3x(zmax) zmax at the nodes for the situation u2z(X = L) < \u03b4 and u2z(X = L) \u2265 \u03b4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000478_chicc.2015.7260360-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000478_chicc.2015.7260360-Figure5-1.png",
        "caption": "Fig. 5: The polar coordinate of a straight line.",
        "texts": [
            " Through looking for the local maximum value in accumulator results, the task for detecting arbitrary shape is transformed into the problem of peak statistics. The busbar has linear structure and the busbars are approximately parallel to each in an image. Hence we use the standard Hough Transform to detection the straight lines. A line in 2-dimensional space using the parametric representation is as follow \u03c1 = x \u00b7 cos\u03b8 + y \u00b7 sin\u03b8 (7) where the variable \u03c1 is the perpendicular distance of the origin to the line along a perpendicular vector. \u03b8 is the angle between this vector and the positive x axis (shown in Fig. 5). Thus, the Hough Transform provides a mapping function (x, y) \u2192 (\u03c1, \u03b8). Then the Hough Transform generates a parameter space matrix whose rows and columns correspond to \u03c1 and \u03b8 values, respectively. In clustering problem, the sample set is denoted as S = {x(1), x(2), \u00b7 \u00b7 \u00b7 , x(m)}, x(i) \u2208 Rn. K-means is one of the simplest clustering algorithm [13]. It minimize the error sum squares of each sample to the mean value of corresponding class. The conventional K-means is described as \u2022 Randomly select k(k > 0) cluster centroids as \u03bc1, \u03bc2, \u00b7 \u00b7 \u00b7 , \u03bck \u2208 Rn"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002892_b978-0-12-801950-4.00006-8-Figure6.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002892_b978-0-12-801950-4.00006-8-Figure6.4-1.png",
        "caption": "FIG. 6.4",
        "texts": [
            " , N, are both sufficiently small, it is easy to verify that the dynamics can be linearized to the following form: { \u03b4\u0307i(t) = fi(t), Mi f\u0307i(t) + Di fi(t) = Pi m(t) \u2212 \u2211 k\u223ci cik( \u03b4i \u2212 \u03b4k), (6.14) where Pi m is the difference between the mechanical power and the stable one, and cik = V2Rik |Zik|2 sin \u03b4ik \u2212 V2Xik |Zik|2 cos \u03b4ik, (6.15) where \u03b4ik is the stable phase difference between generators i and k in the equilibrium state, Zij = Rij + jXij is the impedance of the transmission line between generators i and j, Ei is the voltage of generator i, and Yi is the shunt admittance. The derivation procedure can be found in Ref. [100]. The power network node is illustrated in Fig. 6.4. To facilitate the analysis, we make the following assumptions, which further simplify the dynamics of the power grid: \u2022 The damping constant is identical for all generators, and is denoted by D. \u2022 The phase difference \u03b4ik is relatively small. Hence we can assume cik = \u2212V2Xik |Zik|2 , (6.16) which implies cik = cki. \u2022 All the rotor inertias are the same and equal to 1, i.e., Mn = 1. It is easy to extend this to the general case. For notational simplicity, we have cik = 0 if generators i and k are not adjacent"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure7.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure7.8-1.png",
        "caption": "Fig. 7.8 Different modeling approaches for the beam shown in Fig. 7.7: a top load, b bottom load, and c equally distributed load",
        "texts": [
            " The analytical solution is obtained as uz,max = \u2212 32F Ea for the Euler\u2013Bernoulli beam and as uz,max = \u2212 32F Ea \u2212 24F(1+\u03bd) 5Ea for the Timoshenko beam. 7.2 Advanced Example: Different 3D modeling approaches of a simply supported beam Given is a simply supported beam as indicated in Fig. 7.7. The length of the beam is 4a and the square cross section has the dimensions 2a \u00d7 2a. The beam is loaded by a single force F acting in the middle of the beam. Note that the problem is not symmetric. 336 7 Three-Dimensional Elements (a) 7.2 Finite Element Solution 337 modelling approach is based on two elements with nodes 1, . . . , 12, see Fig. 7.8, and different ways of introducing the acting force F . The solution should be given as a function of F, a, E, \u03bd = 0.3. 7.2 Solution The elemental stiffnessmatrix of an eight-node hexahedronwith dimensions a\u00d7a\u00d7a is known from problem 7.1. The dimension of this matrix is 12 \u00d7 12. Considering two solid elements, a global system of equations with the dimension 36 \u00d7 36 must be assembled. Introducing the support conditions, i.e. u1X = u1Y = u1Z = u4X = u4Y = u4Z = 0 and u9Z = u10Z = 0 reduces this system to a dimension of 28\u00d7 28"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001399_1464419315571983-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001399_1464419315571983-Figure1-1.png",
        "caption": "Figure 1. Structure of a rotor-coaxial water pump bearing system.",
        "texts": [
            " The VC effects of the bearing are described as periodically repeated excitations. The force variation and impact formation when the rolling elements roll over a localized race defect is simulated by the pulse excitation of finite width. The influences of dynamic bearing loads caused by the rotor unbalances on the VC excitation and the defect excitation are both taken into consideration. System structure and loading analysis The common structure of a rotor-water pump bearing system with a cylindrical roller row and a ball row is shown in Figure 1. On the spindle of the bearing are installed the pump impeller, driven pulley, and cooling fan. The thin outer ring of the bearing is mounted with an interference fit in the aluminum pump shell housing. The housing has several pairs of strengthening ribs around its outer diameter to obtain a large stiffness, which can reduce the outer-ring bending or ovalization. The external static loads acting on the spindle of the bearing include belt tensions from the driven pulley, axial forces from the pump impeller, and cooling fan and the gravity of each component installed on the bearing spindle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003759_gt2016-56951-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003759_gt2016-56951-Figure4-1.png",
        "caption": "FIGURE 4. NEW ROTOR DESIGN",
        "texts": [
            " Compared to the former design, it was necessary to re-design the thrust bearing and the displacement device. The displacement device is needed to adjust the axial rotor position. The new setup includes an extended shaft which leads through the thrust bearing. So, a telemetry system can be installed at the rotor shaft end. One main aspect for the development of the new rotor system was the capability of interchangeable running surfaces. This requirement is realized using integrated rotor adapters in the form of annular rings. Figure 4(a) shows a 3D-Model of the new rotor system design. The dimensions of all relevant diameters are identical compared to the former solid rotor body concept, in order to enable a direct comparison of future results with previous measurements. The geometrical dimensions of each adapter are identical except for the staggered surface at the drive end (flow guides). The shaft and both adapters are made of a standard CrMoV alloy, typically used in steam turbines. The high-loaded screw connections have been designed and calculated according to German VDI 2230 guideline. The inner geometry of the new rotor is depicted by a sectional view in Figure 4(b), relating to the cutting planes in Figure 4(a). The instrumentation of thermocouples in the rotating system requires a hollow shaft design in order to rout the thermocouple wiring. Superficial cable guidance is impossible due to the complex rig design. Starting from the open rotor at the telemetry system, the rotor has a centered bore hole with nearly half of the total rotor length. Two opposing, inclined bore holes build the way out of the system at the drive end (view A-A). Displaced by 90\u25e6 (view B-B), two additional bore holes are implemented to lead instrumentation wires to the displacement end",
            " The whole development process was supported by FEM calculations in order to guarantee a safe test operation. Despite the 5 Copyright \u00a9 2016 by Siemens Energy, Inc. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 09/03/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use high thermal and centrifugal loads, material stresses and stability are uncritical for the described rotor design. The new concept includes an interference section between each adapter and the shaft body in order to avoid lifting the adapters from the shaft surface caused by centrifugal forces (c.f. Figure 4(b)). Figure 7 verifies the uniform radial growth of the combined shaft-adapters system. The appropriate calculations were conducted for a nominal rotor speed of 10.000 rpm and an ambient temperature of 500 \u25e6C. Brush seal impacts by means of local heat influxes due to frictional contacts have not been considered within the design calculations. In contrast to previous rotor designs [25] centrifugal lifting effects of the adapters, influencing the clearance between rotor and bristle pack could be excluded"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure12-1.png",
        "caption": "Figure 12 Proposed locking mechanism (see online version for colours)",
        "texts": [
            " As shown in Figure 11, the two manipulators share one rail that is installed at the centre. The rail is fixed to one manipulator, while the pulleys to guide the rail are fixed to the other manipulator. By this mechanism, the centre of rotation of each joint corresponds to the centre of the manipulator. Thus, problem P1 in Section 4 is solved. In addition, the size of the duplex system is reduced from that of the conventional one. The locking mechanism was also improved to realise the semi-circular shape in Figure 12. By pushing the piston attached to the end of the manipulator, the hose is expanded by air, and the pins engage to the holes of the link. This locks the joint. To solve P2, a pulling mechanism was employed. Figure 13 shows the pulling mechanism. Both manipulators have the same mechanism, but only one is shown in order to simplify the illustration. As shown in Figure 14, by pulling the wire of the locked manipulator, the movable manipulator is pulled forward. This reduces the friction, and the duplex manipulator can escape from its deadlock"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000464_b978-0-08-096532-1.00508-2-Figure18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000464_b978-0-08-096532-1.00508-2-Figure18-1.png",
        "caption": "Figure 18 Illustration of single crystal spiral grain selector with design parameters. Reproduced from Dai, H. J.; Gebelin, J. C.; Newell, M.; Reed, R. C.; D\u2019Souza, N.; Brown, P. D.; Dong, H. B. Grain Selection during Solidification in Spiral Grain Selector. In Superalloys 2008; Reed, R. C., et al., Eds.; TMS: Champion, PA, 2008; pp 367\u2013374.",
        "texts": [
            " Different single crystal (SX) selector designs have been employed (Figure 17) (26,67) to select a single grain. The restrictor selector was found to be less efficient since a longer length is required to select a single grain. In the angled selector, problems occur around the corners because new grains often nucleate due to the sudden change in growth direction and the high temperature gradient (33,68). The most commonly used grain selector is the spiral selector. The spiral grain selector consists of two parts (Figure 18): a starter block referring to competitive growth for the grain orientation optimization and a spiral grain selector facilitating dendrite branching to ensure that only single grain eventually survives at the top of the seed. The parameters used to define the geometry of the spiral grain selector include the diameter of starter block (dB), the height of the starter block (LB), spiral takeoff angle (q), corresponding pitch length (LP), spiral diameter (dS), and wax wire diameter (dW) (Figure 18). Using a Bridgman furnace, SX components (e.g., SX turbine blade) can be produced by withdrawing the mold to allow directional solidification. During casting, a number of grains nucleated within the starter block can grow into the spiral grain selector. Invariably only one grain survives after the competitive growth in the spiral grain selector, growing into the main body of the component as a single crystal with further aid of controlled processing parameters (e.g., high thermal gradient and low pulling velocity)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001174_0954405414554016-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001174_0954405414554016-Figure12-1.png",
        "caption": "Figure 12. Transmission system schematic diagram of an antibacklash gear angle and angle-contact ball bearing.",
        "texts": [
            " The combination of the dual mapping results ((1) and (2)) established a relationship (3) h= f g between alignment process parameters and the dynamic characteristics of the assembly (in Figure 11). We can change the process parameters to satisfy the dynamic demand of the assembly and make the alignment decision to optimize the process scheme and mining assembly data to predict the assembly performance. The transmission system model of an anti-backlash gear angle and angle-contact ball bearing can be simplified as shown in Figure 12(a). The outer ring of the bearing is a fixed and non-massive auxiliary ring is added with the same angular velocity as a motion unit M=M1 +M2 +M3, where the bearing inner ring is represented by M1, the transmission shaft is denoted by M2, and the anti-backlash gear is given as M3. Under the action of bearing contact stiffness and anti-backlash gear torsional stiffness, the motion unit presents 4 degrees of freedom: rotation around the shaft ku, axial translation along the axis of shaft ka, and two radial translations perpendicular to the axis of shaft kr, k 0 r (Figure 12(c)). The rolling bodies are constrained by springs linking the auxiliary outer ring and motion unit which guarantee an adequate position of the rolling bodies. Using the dynamic simulation software program ADAMS, we can simulate the frequency response of the transmission system. As shown in Table 5, some motion pairs are added to ensure the system\u2019s 4 degrees of freedom (as given in Figure 12(d)). In Figure 12(c), the torsional stiffness ku is transformed by the contact stiffness of the anti-backlash gear ku = kst3R2 2 \u00f034\u00de where kst represents the anti-backlash gear contact stiffness and R2 is the base radius of the anti-backlash gear. The axial stiffness ka represents the synthetic axial stiffness formed by all springs in parallel ka = XNb j=1 kcd 1=2 sin2 (a) \u00f035\u00de The radial stiffness kr, k 0 r represents the synthetic radial stiffness formed by all springs in parallel kr = XZ 1 kcd 1=2 cos2 a\u00f0 \u00de cos2 cj k0r = XZ 1 kcd 1=2 cos2 a\u00f0 \u00de sin2 cj In ADAMS software, the rolling body is regarded as a spring, d represents the deformation of the spring at NANYANG TECH UNIV LIBRARY on June 5, 2016pib"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002508_icra.2016.7487792-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002508_icra.2016.7487792-Figure2-1.png",
        "caption": "Fig. 2. Planar Serial 3-DOF Manipulator.",
        "texts": [
            " Through the results of Theorem 3, we see that the PD controller Kpd of (13) with the Conditions i)\u2013iii) mathematically guarantees ISS of the system (9) and the L1 performance (12). Furthermore, this controller would make the L\u221e norm of the composite error s small, and this will be experimentally demonstrated in the following section. This section examines the effectiveness of the method for designing the L1 optimal PD controller in reducing the L\u221e norm \u2016s\u2016\u221e through experiments. To this end, we employ a planar serial 3-degrees of freedom (DOF) manipulator as shown in Fig. 2 [32]. This 3-DOF manipulator is influenced by the gravity and consists of three brush-less direct current (BLDC) motors and three linkages whose lengths are 0.3 m, 0.3 m and 0.1 m, respectively. The motor of Joint 1 has 19.3 Nm capability and those of Joint 2 and Joint 3 have 13.5 Nm capabilities, respectively. This robot system is operated with the sampling frequency 1 kHz in a synchronous fashion. We consider the desired trajectories as shown in Fig. 3, which are sixth-order polynomial functions of time so that the joint angles and joint velocities can be set to be zeros when t = 0 s, 5 s and 10 s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001399_1464419315571983-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001399_1464419315571983-Figure3-1.png",
        "caption": "Figure 3. Load analysis of a single shaft and disc.",
        "texts": [
            " In order to conduct vibration analysis for the rotor-water pump bearing system, the continuous rotor is discretized using the lumped parameter method into n rigid thin discs and n 1 elastic massless shafts as shown in Figure 2. Each rolling element row of the bearing is modeled as a set of springs and dampers. The axial positions located by the center cross sections of roller row and ball row of the bearing are referred to as \u2018\u2018position R\u2019\u2019 and \u2018\u2018position B,\u2019\u2019 respectively. In the fixed coordinate system defined in Figure 3, the x, y, z axes denote horizontal direction, vertical direction, and axial direction of at Monash University on September 18, 2015pik.sagepub.comDownloaded from the rotor, respectively. The origin of coordinates is located at the center of the first disc, i.e. the fan center. As shown in Figure 2, the bearing housing including the outer ring is simplified as a rigid body with rotary inertia supported by a fixed pivot and an equivalent spring. This means that the housing can deflect around the fixed pivot. The equivalent spring for the housing with certain stiffness is supposed to act at the axial position of the rolling element row away from the pivot and the housing ovalization is ignored. The loads acting on individual discs and shafts in the discrete rotor system are illustrated in Figure 3. The dynamic loads on a disc include a shear force increment, F , and a bending moment increment, M, besides the shear force, F, and bending moment, M, transmitted from the shaft at each disc end. The shear force increment includes elastic restoring force, damping force, inertia force, and external excitation force. The bending moment increment includes elastic restoring moment, inertia moment, gyroscopic moment, and external excitation moment. Usually, the stiffness and damping of two rolling element rows of the water pump bearing are anisotropic under the external loads",
            " Because the bearing housing has been simplified as a rigid body deflecting the pivot angularly, its motion equations can be expressed as Pn j\u00bc1 Cty \u00f0 _yh _y\u00deLo \u00fe Kty \u00f0 yh y\u00deLo K tx \u00f0 hx x \u00de j Jhx \u20ac hx Khy hxL 2 h \u00bc 0 Pn j\u00bc1 Ctx \u00f0 _xh _x\u00deLo \u00fe Ktx \u00f0xh x\u00deLo K ty \u00f0 hy y \u00de j Jhy \u20ac hy Khx hyL 2 h \u00bc 0 xhj \u00bc hyLo j, yhj \u00bc hxLoj 8>>>>>>>>>>>< >>>>>>>>>>>: \u00f02\u00de In order to take the effect of shear deformation on total deformation of the shaft into consideration, a parameter is defined and introduced into equation (1). \u00bc 6 EI G0Asl2 \u00f03\u00de where is a factor to describe the distribution of shear stress on the cross section of the shaft. If a harmonic forced vibration is generated in the rotor-bearing system, the state variable of any disc can be given in complex field by Zj \u00bc Z\u0302j e i! t \u00f04\u00de Then, regarding for the force and moment directions defined in the coordinate system shown in Figure 3, the force and moment increments of disc j can be written as My Mx Fx Fy 0 BBB@ 1 CCCA j \u00bc Jy! 2 y K ty\u00f0 y hy\u00de iJz!s! x \u00fe pmey Jx! 2 x K tx\u00f0 x hx\u00de\u00feiJz!s! y \u00fe pmex m!2x\u00fe iCtx!\u00f0x xh \u00de \u00fe Ktx\u00f0x xh\u00de pfex m!2y\u00fe iCty!\u00f0 y yh \u00de \u00fe Kty\u00f0 y yh \u00de pfey 0 BBB@ 1 CCCA j \u00f05\u00de As shown in equation (5), the force and moment increments of any disc have been expressed by the external excitations on the disc and the displacements of the disc and housing. After the stiffness as well as damping for each disc and the stiffness for the housing are all determined, the complex amplitudes of the state variables at one rotor end and the complex amplitudes of the housing angular displacements could be directly searched by the numerical method (e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001724_ssd.2015.7348203-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001724_ssd.2015.7348203-Figure7-1.png",
        "caption": "Fig. 7. Schematic model of mobile robot",
        "texts": [
            " After that, the best beam is calculated by maximizing the following objective function : f(\u03c11, \u03c12, d) = \u03b1 ( d cos(|\u03b5|) dmax ) \u2212 \u03b2 \u2223\u2223\u2223 \u03b5 \u03c0 \u2223\u2223\u2223 (6) where \u03b1 and \u03b2 are weight constants adjusted by experimentation with \u03b1+\u03b2 = 1 ; d cos(|\u03b5|)/dmax is the projected distance over the goal direction where \u03b5 is the angle between the goal direction and the robot\u2019s current orientation, defined as \u03b5 = \u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 0 if \u03c11 \u2264 \u03c10 \u2264 \u03c12 \u03c10 \u2212 \u03c12 if \u03c12 < \u03c10 \u03c11 \u2212 \u03c10 otherwise (7) where \u03c10 is the goal direction. The beam that maximizes the objective function is selected as the best beam. To determine the desired goal heading, we adopt the goal heading calculation method presented in [13]. After all that, CVM will try to follow this heading, so the robot would not pass excessively close to the obstacles. The schematic model of the mobile robot is shown in Fig.7. The kinematic model is given by: \u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 X\u0307R = VR+VL 2 cos\u03b1R Y\u0307R = VR+VL 2 sin\u03b1R \u03b1\u0307R = VR\u2212VL L (8) The proposed problem in this paper is to allow the robot to the reliable trajectory in an environment with obstacles and take into consideration the physical constraints of the robot. Beam Curvature Method and Dynamic Windows Approach (DWA) are considered among the techniques used in obstacles avoidance problems. Indeed, these strategies not only achieve the obstacle avoidance but also ensure to reach the goal with a high speed navigation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000821_amm.756.85-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000821_amm.756.85-Figure4-1.png",
        "caption": "Fig. 4 - Scheme of measuring of the dynamic characteristics of the electric generator of mechanical pulses: 1 - body, 2 - crank shaft, 3 - rod, 4 - piston-rod 5 - working pneumatic chamber, 6 - hammer head 7 - wave guide 8 - axle-box, 9 - triangulation laser sensor, 10 - power dog, 11 - stand, 12 - direct the laser beam 13 - reflected laser beam",
        "texts": [
            " The merit of usage for of RPE for SPD hardening of such devices is initially laid in design the independent regulation of energy and shocks frequency, small size, relatively low cost (hydraulic hammers), relatively (pneumatic hammers) high efficiency (40 ... 50%), lack of need to use additional devices of electric power supply of drive, such as oil pumping station for hydraulic hammers or compressor for pneumatic hammers. The operating principle of the electro-pneumatic hammer is in the following. Electric motor, mounted in the housing 1 (fig. 4) rotates the crank shaft 2. At that piston-rod 3, connected to the crank, transmits reciprocating motion to 4. During upstroke, in pneumatic chamber 5 of hammer head 6 below the piston the vacuum occurs. As a result of it, the hammer head starts to move up following piston, cocking for speedup. Then the piston, passing the top dead center, begins to move down against the hammer head, and the air, trapped between the hammer head and the piston begins to contract. Hammer head stops, and then starts moving down (operating stroke)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure13-1.png",
        "caption": "Fig 13 Eddy current density distribution",
        "texts": [
            "The effect caused by eccentric angle is also very small. But as Fig 12 shows the different eccentric angle will affect the distribution of rotor loss. In a healthy machine, the eddy current loss in each magnet is equal. When the dynamic eccentric angle =0\u00b0, PM1 has the minimum air gap length, which leads to the maximum harmonic magnetic field making eddy current loss increase. PM3 has the maximum air gap length, so the eddy current loss on PM3 is the smallest. PM2 and PM4 are symmetric, so they have the same eddy current loss. Fig 13 gives Fig 14 shows that the rotor loss at different eccentricity ratio when eccentricity angle is zero. It is seen that dynamic eccentricity has a small effect on rotor loss. But the loss will focus on PM1 which will make it produce larger temperature rise. Therefore, we should try to avoid dynamic eccentricity. IV. CONCLUSION Base on the eccentric air gap magnetic field, the analytical expression of eddy current loss incurred on the rotor magnet under no-load conditions is derived. The relationship between additional eddy loss and eccentricity ratio is presented"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003477_pedes.2016.7914438-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003477_pedes.2016.7914438-Figure1-1.png",
        "caption": "Fig. 1. Principle sketch of the stator cage machine.",
        "texts": [
            " Ampere\u2019s Law then leads to the magnetic field strength and consequently the inductance will be calculated evaluating the magnetic energy. The analytical calculation method is validated by diverse FEM simulations. Keywords \u2014 Stator Cage Machine; ISCAD; analytical calculation; inductance, magnetic energy I. INTRODUCTION Recently, a new kind of electrical machine, so-called \u201cStator Cage Machine\u201d, has been developed [1]. The stator consists of a stack of iron lamination (like for conventional electrical machines), but with massive conductors in each slot, being short-circuited at one axial end of the machine, see Fig. 1. At the other axial end of the machine, each slot conductor is connected to the center tap of a half bridge, so that the current in each slot can be determined individually. The only boundary condition for the stator slot currents is that the sum of all stator slot currents is equal to zero. This kind of electrical machine together with the driving power electronics has been called ISCAD (Intelligent Stator Cage Drive) [1]. Using a squirrel-cage rotor, it is possible to change the number of pole pairs during operation of the machine"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001062_s00707-013-1041-9-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001062_s00707-013-1041-9-Figure1-1.png",
        "caption": "Fig. 1 Illustration of transmission system with lightened discs of wheels in aerospace application",
        "texts": [
            " The research is motivated primarily by the whole series of pending issues in the area of dynamic phenomena that influence and closely relate to the safety and dependability of such systems. The origin and existence of these phenomena are not still wholly known in terms of both theoretical and experimental investigation in many cases of such complicated systems. The lightening holes in discs of the cog wheels often occur at light high-speed drive systems, for example, at the speed reducers of turbines onto propellers, see Fig. 1. These holes cause the speed heteronomy in the damping properties of cog wheel discs. Generally, the lightening disc holes as well as the forms of discs can have the different forms according to the functional requirements. Consequently, the radial stiffnesses\u2013masses and the damping characteristics that result of these properties, can have the complicated functional relations that depend on the revolutions of appropriate wheels. The paper is organized as follows: Sect. 2 describes the analysed mathematical\u2013physical model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002042_arso.2015.7428209-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002042_arso.2015.7428209-Figure5-1.png",
        "caption": "Fig. 5. Case Study: the OpenROV with the Schonflies motion space is to position itself with the correct position and angle with respect to the object to be inspected.",
        "texts": [
            " In both cases we find that using the standard interface of the OpenROV Software where the robot is controlled using the keyboard has room for improvement. In this section we discuss how our control algorithm can be used to improve the performance of the operation for different kinds of motion. Consider the case where the operator is to control an underatuated ROV in the Schonflies motion group. The operator wants the camera to point at an object that lies above the ROV, and with the camera pointing at the object with a specific angle, for example horizontally. The scenario is illustrated in Figure 5. To obtain this, the operator needs to control both the elevation (z-direction) of the robot and the rotation around the y-axis which is the orientation with respect to the horizontal plane (\u03b8-direction). The first is controlled using the thrusters while the second is controlled by tilting the camera itself. If we ask different operators to perform this operation, they will intuitively lift the haptic device to elevate the robot, and at the same time they will lower the haptic tip to change the orientation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002495_s1068798x16010159-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002495_s1068798x16010159-Figure4-1.png",
        "caption": "Fig. 4. Calculation of the threaded joint by the method in [2] (a) and in the present work (b).",
        "texts": [
            " (3) When the samples are first loaded, \u03b4 is greater than in subsequent loading (curves 2 and 3), on account of plastic deformation of the surface micro projections at the first contact. (4) After loading for the second or third time, the deformation of the microprojections becomes elastic. The operation of a threaded joint characterized by contact pliability of the tightening screws was consid ered in [2]. On that basis, a system of z + 3 equations was written for a threaded joint with a rectangular flange tightened by z screws (with a tightening force Fti for each screw), under the action of tipping torque M (Fig. 4a). The system of equations consists of the equi librium conditions for the forces and torques, the equations for the strain of the bracing flanges and the contact layer between them, and the equations of screw deformation (1) Here Fsi is the force in screw i; b is the width of the flange; px is the pressure at the contact surface of the flanges, a distance x from the neutral axis; xi is the distabce of screw i to the neutral axis; t is the dis tance from the neutral axis to the center of rotation under the action of the screws\u2019 tightening forces and torque M; \u03b8 is the angle of rotation of the joint; \u03bbf = (h1 + h2)/(AE) is the pliability of the flanges; h1 and h2 are their thickness values; A is their contact surface area; \u03bbs is the screw pliability [2]; \u03b5 is the scale factor [2]; c is the contact rigidity; \u03b4ti = + \u03bbfApti is the contact increment of the parts under the action of the tightening forces; pti = zFti/A is the pressure at the con tact surface of the flanges created by the tightening forces",
            " (1), we may determine the displace ment x0 of the skew axis due to the nonlinear pressure dependence of the contact increment (3) We use the experimental apparatus in Fig. 5 to determine the position of the skew axis when the threaded joint is subject to a tipping torque. It consists of a split beam with a spacing of 980 mm between the supports. The beam is compressed by means of a threaded joint, which has two rectangular Duralumin flanges (E1 = E2 = 0.7 \u00d7 105 MPa) and four M8 screws (strength class 8.8). The dimensions of the flange\u2019s supporting surface are a = 114 mm, b = 52 mm (Fig. 4); the thickness values are h1 = 30 mm and h2 = 40 mm; and the distance from the screws to the neutral axis is x = 44 mm. The contact surfaces undergo circu lar planing by means of a single tooth mill. According to measurements by a TR220 profilometer at different points of the contact surface, the roughness Ra1 = Ra2 = 11 \u03bcm. The screws are tightened by means of a dynamometric key (torque 10 N m). With a frictional coefficient f = 0.1 in the thread and at the end, that corresponds to a tightening force Fti = 10 kN",
            " In practice, threaded joints are often subjected to a tipping torque that varies according to a symmetric cycle. In the present study, that is simulated by 180\u00b0 rotation of the joint around the neutral axis before the next measurement. The test results show that, in this case, the displacement of the skew axis relative to the neutral axis is the same as in the absence of rotation. Therefore, on rotation of the tipping torque, the skew axis is tangential to an ellipse described by the coordi nate x0 (Fig. 4b). Experiments show that the displacement of the skew axis relative to the neutral axis is 18.8 mm. Com parison with the calculated displacement x0 of the skew axis shows that, when Eqs. (1)\u2013(3) are used, the error is less than for the methods recommended in [1, 4]. Hence, the displacement x0 of the skew axis under the action of a tipping torque is determined by Eq. (2). Thus, these findings serve as further confirmation that calculations of a threaded joint with flexure resis tant flanges, when the surface roughness of the contact surfaces is Ra > 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.16-1.png",
        "caption": "Fig. 35.16 Spherical roller bearing.",
        "texts": [],
        "surrounding_texts": [
            "In a taper roller bearing the line of action of the resultant load through the rollers forms an angle with the bearing axis. Taper roller bearings are therefore particularly suitable for carrying combined radial and axial loads. The bearings are of separable design, i.e. the outer ring (cup) and the inner ring with cage and roller assembly (cone) may be mounted separately. Single row taper roller bearings can carry axial loads in one direction only. A radial load imposed on the bearing gives rise to an induced axial load which must be counteracted and the bearing is therefore generally adjusted against a second bearing. Two and four row taper roller bearings are also made for applications such as rolling mills."
        ]
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.12-1.png",
        "caption": "Fig. 3.12 Five-link wheel suspension (Raumlenkerachse)",
        "texts": [
            " The corresponding relations to Eqs. (3.5) and (3.7) then are f \u00bc 3 nB nG\u00f0 \u00de \u00fe XnG i\u00bc1 fGi ; \u00f03:8\u00de f \u00bc XnG i\u00bc1 fGi 3nL: \u00f03:9\u00de Two independent wheel suspensions that are widely used in automotive engineering are presented subsequently as examples for closed kinematic chains with different topological complexities. The five-link suspension The five-link suspension (in German \u2018\u2018Raumlenkerachse\u2019\u2019) embodies a closed kinematic chain with a very high degree of inner coupling of the inherent kinematic loops (Fig. 3.12). Because in the assignment of degrees of freedom of the system, the isolated rotation of the suspension arms around their longitudinal axes are neglected, the spherical joints at the chassis end are modeled as CARDAN joints, without loss of generality. Double wishbone wheel suspension In contrary to the five-link wheel suspension, the double wishbone suspension (Fig. 3.13) possesses only a weak kinematic coupling between the inherent two kinematic loops. Thus the associated kinematic analysis can be stated and solved explicitly, as shown in Sect"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000395_978-94-007-4132-4_8-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000395_978-94-007-4132-4_8-Figure6-1.png",
        "caption": "Fig. 6. Hoecken\u00b4s six-link dwell mechanism as a model mounted on a Reuleaux frame",
        "texts": [
            " In the beginning of this booklet Hain dealt with the systematics of dwell mechanisms based on six links of the type \u201cWatt\u201d and \u201cStephenson\u201d and found nine structures, Fig. 5. The driving link is marked by a circle segment with one arrow, the driven link with the possibility to have a dwell position by a circle segment with two arrows. Turning pairs with a full rotation of 360 deg are marked by a double circle, those only swinging are marked by a simple circle. So, when we bear in mind Hain\u00b4s systematic overview and look at Hoecken\u00b4s six-link dwell mechanism (Fig. 6), we find out at once that Hoecken\u00b4s choice corresponds with the number 1 of Hain\u00b4s catalogue: there is a four-link crank-rocker with a two-bar E-F coupled to it. It is very interesting to discover that Hoecken put his linkage onto a typical Reuleaux frame and assigned to it the model number 2201. There is also a paper of Hoecken on dwell mechanisms [8], but it does not describe his model shown in Fig. 6. Before manufacturing the model \u201csix-link dwell mechanism\u201d, Hoecken made a technical drawing of it on February 11, 1932 and named it \u201cRastgetriebe\u201d, Fig. 7. The drawing shows details of the links, its dimensions and even a counterweight on the opposite end of the rocker for static balancing. The dimensions of Hoecken\u00b4s six-link dwell mechanism are as follows: n0 = B0C0 \u2248 124.0 A0C0 \u2248 215.0 n1 = A0B0 \u2248 127.0 n2 = A0A = 61.3 n3 = AB = 131.9 n4 = B0B = 94.2 n5 = CP3k = 83.9 n6 = C0C = 139.5 AP3k = 84"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.3-1.png",
        "caption": "Figure 9.3 Laminated rotor configurations of PM machines with embedded or rotor-surface magnets include (a) embedded magnets, (b) inset magnets, (c) I-type, (d) embedded magnets with laminated pole shoes and magnet pockets, (e) flux concentrating version of version (d), and (f) synchronous reluctance machine with rotor surface magnets on d-axis. Constructions (a) through (e) provide saliency, and Lq> Ld. With construction (f), Ld>Lq.",
        "texts": [
            " There are a number of alternative laminated rotor configurations available to achieve the desired PMSM performance characteristics. The PMs can be glued to the rotor outer diameter surface as is done with a solid rotor yoke. However, the magnets can also be embedded, in part or completely. Using the embedded magnet approach, the magnets can be mounted in different positions and orientations. Some basic configurations for laminated PM rotor structures producing different machine properties illustrated in Figure 9.3. For the configurations illustrated by Figure 9.3a\u20139.3c, the physical air gap is approxi mately constant, and the PMs produce an approximately trapezoidal air-gap flux density. Depending on winding arrangement, the voltage induced in the stator of such a machine may include harmonics. These harmonics may affect the torque output of the machine resulting in vibration and noise. Since smooth torque production is normally required, the stator and rotor current linkages and the fundamental must not include same order harmonics. This would result in harmonics-generated torque components or torque ripple",
            " In these machines, some torque is normally produced by the reluctance differences, that is, the difference in the direct and quadrature inductance directions. In these machines, the q-axis inductance Lq is larger than the d-axis inductance Ld. The addition of magnets in these hybrid machines makes their behavioural characteristics considerably better than the behavioural characteristics of a synchronous reluctance machine. In particular, efficiency and power factor can be notably improved. The rectangular magnet configuration shown in Figure 9.3c requires flux barriers near the rotor shaft to inhibit through-shaft PM leakage flux. This is a challenging mechanical design problem. Furthermore, the structure offers a good q-axis armature reaction path, which is not always desirable. The configuration does offer, however, large flux density concentration in multiple pole-pair machines. In the Figure 9.3b configuration, the magnets are surface mounted. This construction provides some reluctance difference between the direct and quadrature axes. Because of the reluctance difference, maximum torque is produced at a load angle well above 90\u00b0. In PM machines, maximum torque is often produced at load angles greater than 90\u00b0, since inductance in the q-direction is often slightly higher than in the d-direction. The configurations illustrated in Figures 9.3d and 9.3e were developed to produce smooth and quiet operation at low rotation speeds with a stator with the number of slots per pole and phase 1",
            "3e will also be prone to torque rippling. A surface magnet rotor with sinusoidal flux density should be used in machines with low values of q to produce sinusoidal rotor current linkage with no inductance differences. In all of the first five configurations, Figures 9.3a through 9.3e, the q-axis inductance Lq is larger than the d-axis inductance Ld (that is, Lq>Ld). In each case, the machine must be driven by a current vector that has negative d-axis current. A PMSM with Ld> Lq is also possible. Figure 9.3f illustrates a rotor configuration that combines the characteristics of a synchronous reluctance machine and a rotor surface magnet PMSM. A machine configured in this way should be driven by a current vector with positive d-axis current. The PM-produced machine flux will be strengthened by the armature reaction. The machine designer, in this case, must be careful to avoid PM material hysteresis losses, which can occur if external magnetic field strength varies between positive and negative values. Laminated rotor machines are subject to magnetic flux leakage, which can be reduced by integrating leakage flux guides as shown in Figure 9.3a. The flux guide can be air or some other poor conductor. The poles shown in Figures 9.3d and 9.3e produce a sinusoidally shaped flux pattern that simultaneously reduces magnetic leakage. The utilization of the magnets is still lower in an embedded magnet machine than in a salient pole machine, for example, in which the magnetic flux is almost completely in the air gap. A laminated rotor structure can be used to increase air-gap flux density by using two magnets per pole (Figure 9.3e), in which case the PM cross-sectional area increases in proportion to the machine pole area. More magnetic material is needed to implement this approach, however, so it results in a higher-cost machine. For higher pole-pair numbers p, the I-type configuration (Figure 9.3c) can also be a flux-concentrating construction. Laminated rotor structures with clearly shaped pole shoes (Figures 9.3d and 9.3e) can easily be equipped with damper windings, which fit well into the pole shoes. The q configuration type enables the production of direct online (DOL) machine versions. However, DOL-starting may be difficult in PM machines, and mostly they are designed for variable frequency drives. The pu parameter values for PM machines differ from the pu values of traditional induction machines and synchronous machines in industrial use"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000833_amm.564.422-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000833_amm.564.422-Figure3-1.png",
        "caption": "Fig. 3. Schematic diagram of Vacuum infusion process used to infuse woven Kenaf fabric reinforce with unsaturated polyester.",
        "texts": [
            " This test method used ASTM D3763 Standard Test Methods for High Speed Puncture Properties of Plastics Using Load and Displacement Sensors. The test apparatus used Hydro Shot Shimadzu HITS-P10, High Speed Puncture Impact Testing Machine. Three specimens of pure polyester and each composite were tested for the impact energy. The impact energy, J was converted to impact strength, kJ/m2 [13]. Fabrication of Composite Panel. The untreated and treated composite were prepared using vacuum infusion process (Fig.3). The woven fabric reinforce were laid on a smooth surface which been coated with a mould release agent. The layer of the vacuum infusion system consists of peel ply and distribution media are at above reinforcing fabric and covered with vacuum bag. A vacuum pump was set at -1.0bar and used to compact the reinforcement material and resin distribution. After checking and making sure there are no leakages, the resin inlet is released and the unsaturated polyester resin was infused on the covered area"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002940_icma.2016.7558559-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002940_icma.2016.7558559-Figure1-1.png",
        "caption": "Fig. 1. Cup-shaped diaphragm coupling-shaft system. Under working condition, the torque moment is transferred from the motor shaft to the diaphragm coupling firstly, and then to the turbine shaft, through the interference fit. The geometry of the interference fit before and after joining is given in Fig. 2 a and b,respectively.",
        "texts": [
            " Hence the existing elastic limit solutions of the thickwalled tubes and rotating objects are not accurate for the interference fits when the torques effect is considered. Accordingly, the aim of this paper is to develop a better method to estimate the maximum load capacity of the interference fits with taking account of the torques influence. II. ANALYTICAL MODEL In this study, the structure analyzed here is only the part of the interference fits of the cup-shaped diaphragm coupling- 193978-1-5090-2396-7/16/$31.00 \u00a92016 IEEE shaft system. As shown in Fig.1, there are three components of the cup-shaped diaphragm coupling-shaft system, including a motor shaft, a cup-shaped diaphragm coupling and a turbine shaft. The cup-shaped diaphragm coupling contains two cupshaped cylindrical sides connected by a thin quill shaft. The cylindrical sides fit over the motor shaft and turbine shaft with an interference fit respectively as shown in Fig. 1. The stresses of the coupling and shaft are analyzed firstly in order to derive the elastic limit solutions. Using the classic elastic plane stress theory and the rotating aunular disk shown in Fig.3, the equilibrium differential equation, straindisplacement relation and constitutive equation can be expressed as follows 21 0r rr r r r r \u03b8 \u03b8\u03c4 \u03c3 \u03c3\u03c3 \u03c1\u03c9 \u03b8 \u2202 \u2212\u2202 + + + = \u2202 \u2202 (1) u r\u03b8\u03b5 = (2) r du dr \u03b5 = (3) r r E \u03b8\u03c3 \u03bc\u03c3\u03b5 \u2212= (4) r E \u03b8 \u03b8 \u03c3 \u03bc\u03c3\u03b5 \u2212= (5) where Eq(1) can be derived using the force equilibrium in the radial diaction of the annular plate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002058_embc.2015.7318396-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002058_embc.2015.7318396-Figure2-1.png",
        "caption": "Figure 2. SIRIO tools: needle tool (a), patient tool (b), direction tool (c).",
        "texts": [
            " The device can be used for percutaneous interventions such as biopsy, thermal ablation, percutaneous interventional, localization and positioning of fiducials for radiotherapy in different anatomical districts, such as lungs, bones and kidneys. It reconstructs a 3D model from a data set of acquired CT images through automatic procedures [8]. An example of a 3D SIRIO screen is shown in Figure 1. The hardware components of the system are: the visualization and elaboration unit, the infrared optical sensor, the disposable sterile kit (the needle tool (Figure 2a), the patient tool (Figure 2b), the direction tool (Figure 2c), the posture tracking system, the breathing sensors and the needle support. Following a proper recording of the patient and with a selection of DICOM files of interest, it is possible to see the 3D view of needle and the anatomical region of interest, starting from the reconstructed tomographic scans. The navigation system with virtual reality interfaces shows a real-time projection of the patient\u2019s sections (axial and sagittal) showing the surgical needle. The clinician can visualize the needle advancement inside the patient\u2019s chest through a real-time tracking system, as shown in Figure 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure4-1.png",
        "caption": "Fig. 4 The benchmark of block tolerance model a block structure, b Block benchmark",
        "texts": [
            " DF-1 (front surface of VV front rib) is the reference plane of fourth block installing; DF-2 (central plane of right rib) is the reference plane of VV right rib surface (VV-R-F), which also is the standard of hole position on rib and it is perpendicular with DF-1; DF-3 (imaginary plane) is the perpendicular plane with DF-1 and DF-2 respectively. There are four blocks in the tolerance model. Although each has different shape and position, they have common point in installing, that is contacted with surface of under bracket firstly, then by means of upper bracket side surface and bolts are installed to holes of rib. So the first block is selected to define the reference plane as illustration. Figure 4a is the tolerance model, and Fig. 4b is the first block along with upper bracket and bolts. And Fig. 4b also is the benchmark of first block tolerance model, in which three reference plane will be set up. DF-4 is the bottom contact surface between first block and its under bracket; DF-5 is the side contact surface between first block and its upper bracket, which is required to be perpendicular with DF-4; DF-6 (imaginary plane) is the central plane of central hole on first block and brackets, which is required to be perpendicular plane with DF-4 and DF-5 respectively. The under bracket is assembled to rib of VV by contacted both side surfaces of them, which are fixed by four bolts and four location holes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.63-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.63-1.png",
        "caption": "Fig. 3.63 Honeycomb structure approximated by a regular hexagon: a flat orientation and b pointy or angled orientation",
        "texts": [
            " Calculate the displacement of the point of load application and estimate the macroscopic stiffness Estruct of the frame structure. Simplify your results for the macroscopic stiffness for the special case A = \u03c0d2 4 and I = \u03c0d4 64 , i.e. a circular cross section of the beam elements. The derivation should be performed for the quarter model under consideration of the double symmetry. 3.39 Mechanical properties of idealized honeycomb structure A honeycomb structure should be idealized by a regular hexagon as shown in Fig. 3.63. Such a regular hexagon has all sides of the same length L, and all internal 3.5 Supplementary Problems 183 angles are 120\u25e6. Assume for this simplified approach that the honeycomb structure is represented by a single cell, either in flat or angled orientation. Furthermore, a two-dimensional approach based on a frame structure made of generalized beam elements is to consider. Calculate the displacement of the point of load application and estimate the macroscopic stiffness Estruct of the idealized honeycomb structure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000212_978-981-15-5712-5-Figure19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000212_978-981-15-5712-5-Figure19-1.png",
        "caption": "Fig. 19 Steam locomotive developed by Trevithick in 1801",
        "texts": [
            " This motion is capacitively measured by the second set of electrodes on the periphery. This design, although worked well, could not compete with the resolution of the macromachined hemispherical-shell gyroscopes. One of the reasons for this is that the spokes connecting the ring to the central post distort the mode shapes. The second reason is that the actuation force and the change in capacitance were low. 44 G. K. Ananthasuresh These two limitations were overcome to a large extent with a clever and economical design [7], as explained next. Figure 19a shows concentric rings where adjacent ones are connected to each other in a staggered manner. Compare this with the designs in Figs. 10 and 16a. The design in Fig. 19a inherits the features of these two designs. Consequently, the stiffness of the concentric ring structure is made low, and the distortion at the midpoints of the arc beams is reduced (see Fig. 19b). In Fig. 19c, d, the electrodes for actuation and sensing are shown in top view and isometric views. This design of the micromachined gyroscope achieved resolution better than 0.01 \u00b0/h [29]. It is a great example of economical use of material with minimum etching of material resulting in superior performance. 8 Ubiquity of Electrothermal Microactuators Electrostatic actuators, comb drive actuator being the prime example, are widely used in many MEMS sensors. But the force they generate is rather small\u2014barely moving themselves and only occasionally moving something else",
            " 17 The identified stability curves at two-thirds of tool pass of the bidirectional model for the orders p = pc = pd 0.5 1 1.5 2 2.5 spindle speed [rpm] 104 0 5 10 15 20 25 30 de pt h of c ut [m m ] ref p=1 p=2 p=3 p=4 p=5 p=6 p=7 p=8 p=9 p=10 CTRS Fig. 18 The time-domain simulation of nodal responses at [15000 rpm, 22.5 mm] indicated by star in Fig. 17 0 0.01 0.02 time [s] -400 -300 -200 -100 0 100 200 300 re ge ne ra tiv e di sp la ce m en t [ m ] 0 0.01 0.02 time [s] -1 -0.5 0 0.5 1 re ge ne ra tiv e ve lo ci ty [m /s ] Fig. 19 Stability curves for p = pc = pd compared against the reference at the start of tool pass 0.5 1 1.5 2 2.5 spindle speed [rpm] 104 0 5 10 15 20 25 30 de pt h of c ut [m m ] ref p=11 p=12 p=13 p=14 p=15 Fig. 20 Stability curves for p = pc = pd compared against the reference at two-thirds of tool pass 0.5 1 1.5 2 2.5 spindle speed [rpm] 104 0 5 10 15 20 25 30 de pt h of c ut [m m ] ref p=11 p=12 p=13 p=14 p=15 A method which combines modal truncation and tensor-based general order FDM was developed to suppress all the case-by-case symbolic analyses associated with stability analysis of elastic thin-walledworkpiece"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003251_ijwmc.2016.081160-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003251_ijwmc.2016.081160-Figure1-1.png",
        "caption": "Figure 1 Control system model",
        "texts": [
            " Inspired by the integral separated PID, this paper presents the new method of integral separated PID control; in this algorithm the proportion of items is multiplied by \u03b1 and the integral term is multiplied by \u03b2, at the same time controlling the proportion and integral term to improve the integral separated PID control algorithm. In order to improve the robustness of the system, the control algorithm is inspired by the fuzzy control algorithm, and the error is segmented by different control parameters. Finally the robot can reduce the time of the system to adjust by using the new integral separated PID, and improve the system stability and response speed, and enhance the robustness of the system. In Figure 1, VL is the linear velocity of the left wheel, VR is the linear velocity of the right wheel, VC is the linear velocity of the centre of mass, W is the angular velocity of the centre of mass, L is the distance between two rounds, X and Y are two-dimensional coordinate of the centre of mass of the robot, \u03b8 is the angle of the robot with the x-axis. The posture P and position PC of robot are shown in formula (1).     , , , T c c T C c c P x y P x y     (1) The wheel is considered to be simplex rolling motion in the ground, that is to say the speed of the robot is 0, and the velocity vector of the robot\u2019s centre of mass is shown in formula (2)",
            " 1/ 2 1/ 2 1/ 1/ LC R VV S L L V                 (2) Owing to the existence of a non-singular transformation matrix, the robot\u2019s position and attitude can be changed by controlling the speed of the robot. These two vectors are equivalent in the point of control, and it is not difficult to deduce the robot\u2019s kinematics equation (Zhang et al., 2001; Xu et al., 2015): cos sin 0 1 c c c x v P y                             (3) The robot\u2019s coordinate is obtained by the GPS sensor, and the rotation angle of the robot is measured by the angle sensor (Fang and Shen, 1998; Ni et al., 2015; Stylianou, 2015). According to Figure 1 (b) the coordinates can be introduced to launch target rotation angle \u03b83 of the robot, which is shown in formula (4). 1 0 1 1 0 2 1 2 2 1 3 2 1 tan tan y y x x y y x x                  (4) Discrete PID control algorithm is for the most part expressed as formula (5).            0 1 k p i j d u k k error k k error j T k error k error k T        (5) Integral separated PID control algorithm is shown in formula (6).            0 1 k p i j d u k k error k k error j T k error k error k T         (6) In the above formula, error(k) = yd(k) \u2013 y(k), error(k) is the error between the target rotation angle and the actual rotation angle, yd(k) is target rotation angle, y(k) is actual rotation angle and the output of the controlled object, u(k) is the angle controller and the input of the controlled object"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002957_gt2016-56282-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002957_gt2016-56282-Figure1-1.png",
        "caption": "Fig. 1 Photograph (left) and schematic view (right) of test GFTB.",
        "texts": [
            " The present study extends the experimental analysis in Ref. [14] to measure the dynamic load performance of a test GFTB with base excitations. A single degree-of-freedom model of the rotor supported on a pair of GFTBs in the axial direction identifies the bearing load, stiffness, damping, and loss factor with increasing excitation frequencies, applied dynamic forces, and bearing clearances. The effects of a pad angular offset in the pair of GFTBs on the dynamic characteristics are also studied experimentally. Figure 1 shows a photograph (left) and schematic view (right) of the test GFTB with six pads. Each pad consists of a top foil and a bump foil (strip layer), constructing an elastic support structure, and has a ramp formed by determining the height of the metal block at the leading edge of the top foil to develop hydrodynamic pressure within the film thickness (see Fig. 2 for a schematic view of the foil structure of the pad). Note that the load capacity of GFTBs depends on the compliance of the foil structure and the hydrodynamic pressure generated between the top foil and thrust runner"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.23-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.23-1.png",
        "caption": "FIGURE 6.23",
        "texts": [
            " For the example vehicle used in this text the body was rotated 10 each way. The results for the front end model are plotted in Figure 6.22. The gradient at the origin can be used to obtain the value for roll stiffness used in the equivalent roll stiffness model described earlier. In the absence of an existing vehicle model that can be used for the analysis described in the preceding section, calculations can be performed to estimate the roll stiffness. In reality this will have contributions from the road springs, anti-roll bars and possibly the suspension bushes. Figure 6.23 provides the basis for a calculation of the road spring contribution for the simplified arrangement shown. In this case the inclination of the road springs is ignored and has a separation across the vehicle given by Ls. As the vehicle rolls through an angle 4 the springs on each side are deformed with a displacement, ds, given by \u03b4s = \u03c6 Ls/2 \u00f06:4\u00de The forces generated in the springs Fs produce an equivalent roll moment Ms given by Ms = Fs Ls = ks \u03b4s Ls = ks \u03c6 Ls 2/ 2 \u00f06:5\u00de Calculation of roll stiffness due to road springs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003294_vppc.2016.7791805-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003294_vppc.2016.7791805-Figure4-1.png",
        "caption": "Fig. 4. Stator structure of the unequal tooth machine",
        "texts": [
            " 3 shows the variation of the average output torque for the 10-slot/8-pole equal tooth and unequal tooth SPM machines with the ratio of the third harmonic current to the fundamental one. It can be seen that the unequal tooth machine has higher optimal injected third harmonic current and torque improvement than the equal tooth machine since the unequal tooth machine has higher third harmonic back-EMF, which will be discussed in the following section. IV. ELECTROMAGNETIC PERFORMANCE OF EQUAL AND For five-phase 10-slot/8-pole unequal tooth SPM machine, Fig. 4 shows the stator structure. Assuming that the slots are asymmetric, A is the slot pitch angle for the tooth wounds winding. A can be expressed as: 1 2 1 1 arcsin arcsin 2 2 TW TW A B R R \u2248 + \u2212 (11) where B is the angle between the axes of the neighbouring teeth. For 10-slot machine, B=36 deg. mech. For the five phase 10-slot/8-pole SPM machine, the distribution factor is unity. The winding factor equals to the short pitch factor and can be expressed as: 1 3 sin 2 3 sin 2 W W A K C A K C \u03c0 \u03c0 = = (12) where C is the pole pitch angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003232_detc2016-59194-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003232_detc2016-59194-Figure2-1.png",
        "caption": "Fig. 2 Two configurations of a line symmetric 7R mechanism.",
        "texts": [
            " In order to solve this system with the computer program Maple, one can introduce an extra variable v and add the equation u1u2 \u00b7 \u00b7 \u00b7unv\u22121 = 0 to Eq. (21) and compute an elimination ideal that eliminates v to obtain a new set of equations with the extraneous solutions excluded. SYMMETRIC 7R SPATIAL MECHANISM In this section, the loop equations Eq. (19) in Section 3 will be used for the kinematic analysis of conventional mechanisms. The explicit input-output equations of a line symmetric 7R spatial single-loop mechanism will be derived. A line symmetric 7R spatial single-loop mechanism (Fig. 2) is composed of seven identical links with the following link parameters: di = 0, i = 1, 2, . . . 7 Li = 1, i = 1, 2, . . . 7 \u03b1i = \u03b10, where S\u03b10 6= 0 and i = 1, 2, . . . 7 Since these line symmetric 7R mechanism with different \u03b10 have similar input-output equations, this section will focus on the case with \u03b10 = \u03c0/2. Using Eq. (19), we obtain the following set of kinematic loop equations  e1(t1, t2, \u00b7 \u00b7 \u00b7 t7) = 0 e2(t1, t2, \u00b7 \u00b7 \u00b7 t7) = 0 e3(t1, t2, \u00b7 \u00b7 \u00b7 t7) = 0 g1(t1, t2, \u00b7 \u00b7 \u00b7 t7) = 0 g2(t1, t2, \u00b7 \u00b7 \u00b7 t7) = 0 g3(t1, t2, \u00b7 \u00b7 \u00b7 t7) = 0 (29) where e1 = \u2212t1t2t3t4t5t6 + t1t2t3t4t5t7 \u2212 t1t2t3t4t6t7 + t1t2t3t5t6t7 \u2212 t1t2t4t5t6t7 + t1t3t4t5t6t7 \u2212 t2t3t4t5t6t7 + t1t2t3t4 + t1t2t3t5\u2212 t1t2t3t6\u2212 t1t2t3t7 + t1t2t4t5 + t1t2t4t6\u2212 t1t2t4t7 + t1t2t5t6 + t1t2t5t7 + t1t2t6t7 + t1t3t4t5\u2212 t1t3t4t6\u2212 t1t3t4t7 + t1t3t5t6\u2212 t1t3t5t7 + t1t3t6t7\u2212 t1t4t5t6\u2212 t1t4t5t7\u2212 t1t4t6t7\u2212 t1t5t6t7 + t2t3t4t5 + t2t3t4t6\u2212 t2t3t4t7 + t2t3t5t6 + t2t3t5t7 + t2t3t6t7 + t2t4t5t6\u2212 t2t4t5t7 + t2t4t6t7\u2212 t2t5t6t7 + t3t4t5t6+t3t4t5t7+t3t4t6t7+t3t5t6t7+t4t5t6t7+t1t2\u2212t1t3\u2212 t1t4+t1t5+t1t6\u2212t1t7+t2t3\u2212t2t4\u2212t2t5+t2t6+t2t7+t3t4\u2212 t3t5 \u2212 t3t6 + t3t7 + t4t5 \u2212 t4t6 \u2212 t4t7 + t5t6 \u2212 t5t7 + t6t7 \u2212 1 e2 = \u2212t1t2t3t4t5t6t7 + t1t2t3t4t5 + t1t2t3t4t6 \u2212 t1t2t3t4t7 + t1t2t3t5t6 + t1t2t3t5t7 + t1t2t3t6t7 + t1t2t4t5t6\u2212 t1t2t4t5t7 + t1t2t4t6t7\u2212 t1t2t5t6t7 + t1t3t4t5t6 + t1t3t4t5t7 + t1t3t4t6t7 + t1t3t5t6t7 + t1t4t5t6t7 + t2t3t4t5t6\u2212 t2t3t4t5t7 + t2t3t4t6t7\u2212t2t3t5t6t7+t2t4t5t6t7\u2212t3t4t5t6t7+t1t2t3\u2212t1t2t4\u2212 t1t2t5 + t1t2t6 + t1t2t7 + t1t3t4 \u2212 t1t3t5 \u2212 t1t3t6 + t1t3t7 + t1t4t5 \u2212 t1t4t6 \u2212 t1t4t7 + t1t5t6 \u2212 t1t5t7 + t1t6t7 \u2212 t2t3t4 \u2212 t2t3t5 + t2t3t6 + t2t3t7 \u2212 t2t4t5 \u2212 t2t4t6 + t2t4t7 \u2212 t2t5t6 \u2212 t2t5t7\u2212t2t6t7\u2212t3t4t5+t3t4t6+t3t4t7\u2212t3t5t6+t3t5t7\u2212t3t6t7+ t4t5t6 + t4t5t7 + t4t6t7 + t5t6t7\u2212 t1\u2212 t2 + t3 + t4\u2212 t5\u2212 t6 + t7 e3 = \u2212t1t2t3t4t5t6t7 \u2212 t1t2t3t4t5 + t1t2t3t4t6 + t1t2t3t4t7\u2212 t1t2t3t5t6 + t1t2t3t5t7\u2212 t1t2t3t6t7 + t1t2t4t5t6 + t1t2t4t5t7 + t1t2t4t6t7 + t1t2t5t6t7\u2212 t1t3t4t5t6 + t1t3t4t5t7\u2212 t1t3t4t6t7 + t1t3t5t6t7\u2212 t1t4t5t6t7 + t2t3t4t5t6 + t2t3t4t5t7 + t2t3t4t6t7+t2t3t5t6t7+t2t4t5t6t7+t3t4t5t6t7+t1t2t3+t1t2t4\u2212 t1t2t5 \u2212 t1t2t6 + t1t2t7 + t1t3t4 + t1t3t5 \u2212 t1t3t6 \u2212 t1t3t7 + t1t4t5 + t1t4t6 \u2212 t1t4t7 + t1t5t6 + t1t5t7 + t1t6t7 + t2t3t4 \u2212 t2t3t5 \u2212 t2t3t6 + t2t3t7 + t2t4t5 \u2212 t2t4t6 \u2212 t2t4t7 + t2t5t6 \u2212 t2t5t7+t2t6t7\u2212t3t4t5\u2212t3t4t6+t3t4t7\u2212t3t5t6\u2212t3t5t7\u2212t3t6t7\u2212 t4t5t6 + t4t5t7\u2212 t4t6t7 + t5t6t7 + t1\u2212 t2\u2212 t3 + t4 + t5\u2212 t6\u2212 t7 g1 = \u2212t1t2t3t4t5t6 + t1t2t3t4t5t7 \u2212 t1t2t3t4t6t7 + t1t2t3t5t6t7 \u2212 t1t2t4t5t6t7 + t1t3t4t5t6t7 \u2212 t2t3t4t5t6t7 \u2212 3t1t2t3t4+t1t2t3t5\u2212t1t2t3t6+3t1t2t3t7\u22123t1t2t4t5+t1t2t4t6\u2212 t1t2t4t7\u22123t1t2t5t6+t1t2t5t7\u22123t1t2t6t7+t1t3t4t5\u2212t1t3t4t6+ 3t1t3t4t7+t1t3t5t6\u2212t1t3t5t7+t1t3t6t7\u2212t1t4t5t6+3t1t4t5t7\u2212 t1t4t6t7+3t1t5t6t7\u22123t2t3t4t5+t2t3t4t6\u2212t2t3t4t7\u22123t2t3t5t6+ t2t3t5t7\u22123t2t3t6t7+t2t4t5t6\u2212t2t4t5t7+t2t4t6t7\u2212t2t5t6t7\u2212 3t3t4t5t6 + t3t4t5t7\u22123t3t4t6t7 + t3t5t6t7\u22123t4t5t6t7 +5t1t2 + 3t1t3 \u2212 t1t4 + t1t5 \u2212 3t1t6 \u2212 5t1t7 + 5t2t3 + 3t2t4 \u2212 t2t5 + t2t6 \u2212 3t2t7 + 5t3t4 + 3t3t5 \u2212 t3t6 + t3t7 + 5t4t5 + 3t4t6 \u2212 t4t7 + 5t5t6 + 3t5t7 + 5t6t7 + 7 g2 = \u2212t1t2t3t4t5t6t7 \u2212 3t1t2t3t4t5 + t1t2t3t4t6 \u2212 t1t2t3t4t7\u22123t1t2t3t5t6+t1t2t3t5t7\u22123t1t2t3t6t7+t1t2t4t5t6\u2212 t1t2t4t5t7 + t1t2t4t6t7\u2212 t1t2t5t6t7\u22123t1t3t4t5t6 + t1t3t4t5t7\u2212 3t1t3t4t6t7+t1t3t5t6t7\u22123t1t4t5t6t7+t2t3t4t5t6\u2212t2t3t4t5t7+ t2t3t4t6t7 \u2212 t2t3t5t6t7 + t2t4t5t6t7 \u2212 t3t4t5t6t7 + 5t1t2t3 + 3t1t2t4\u2212t1t2t5+t1t2t6\u22123t1t2t7+5t1t3t4+3t1t3t5\u2212t1t3t6+ t1t3t7+5t1t4t5+3t1t4t6\u2212t1t4t7+5t1t5t6+3t1t5t7+5t1t6t7+ 4 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000821_amm.756.85-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000821_amm.756.85-Figure1-1.png",
        "caption": "Fig. 1 - Scheme of measuring of the dynamic characteristics of the hydraulic generator of mechanical pulses: 1 - hammer head, 2 - wave guide 3 - tool 4 - cocking cavity 5 - accelerating cavity 6 - sleeve, 7 - control valve spool, 8 - hydraulic motor, 9 - throttle, regulating fluid flow in the hydraulic motor, 10 - pressure tank, 11 - rod, 12 - tube, 13 - measuring coil, 14 - oscillograph, 15 - polarizing coil, 16 - polarizing coil power supply , 17 - pressure reducing Valve",
        "texts": [
            " At that it is necessary to conduct researches to determine the kinetic shocks energy of the impact and to ensure its precise regulation. Determination of the kinetic energy for hammers with a known mass of the hammer head shall be come to the determination of the shock velocity. To measure the velocity of the hammer head at its movement during speedup the inductive speed sensors, which transform translational motion of the hammer head to electrical signal, may be used. Such scheme was used for shock mechanism to harden SPD - hydraulic pulse generator (fig. 1), designed on the basis of known structures of hydraulic hammers for mining operations [4]. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications Ltd, www.scientific.net. (#523754121, Link\u00f6pings Universitetsbibliotek, Link\u00f6ping, Sweden-04/01/20,07:04:38) Technical characteristic of the hydraulic pulses generator Shock energy, J\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026\u2026.60...300 Adjustment range of shocks frequency, Hz\u2026\u2026"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000748_ep.12202-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000748_ep.12202-Figure1-1.png",
        "caption": "Figure 1. Schematic diagram of the microbial fuel cell stack. [Color figure can be viewed in the online issue, which is available at wileyonlinelibrary.com.]",
        "texts": [
            " The SPEEK membrane was sandwiched between the prepared anode and cathode electrodes and hot pressed at 708C and 1 ton pressure for 2 min. Thus, the prepared MEA was assembled and sandwiched between the flat graphite plates (anode and cathode) with the cathode side of the MEA facing the air circulation side. All the four units of the stack composed of conductive (Graphite, 10 cm width by 10 cm length by 1.5 cm height) plates that were compressed together, containing a channel that was separated into two portions to form the anode and cathode chambers (Figure 1). By serially aligning the anode, membrane, and cathode (to form MEA) the distance between anode and cathode was reduced. The anode was fluidically connected to each other and the cathode was connected to air. Each anode plate was drilled to form a rectangular channel in a serpentine path (0.5 cm wide and 0.4 cm deep), having a total surface area of 42.5 cm2 and a total volume of 17 cm3, achieving a total Table 1. Membrane properties. Proton conductivity (S cm21) 0.5 3 1022 Ion exchange capacity (meq g21) 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002044_978-3-319-25017-5_28-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002044_978-3-319-25017-5_28-Figure3-1.png",
        "caption": "Fig. 3 Showing how the state and action of the system is defined",
        "texts": [
            " If for example, the selected action is closer to ai,1 than to ai,2, despite its further distance, it replaces ai,2, so as to protect changes range of the action space. The proposed method, at last makes use of a combination of the above mentioned procedure and the \ud835\udf00-greedy method. That is, with probability 1 \u2212 \ud835\udf00, the action is selected with maximum value. But with probability \ud835\udf00, Eq. 9 is used to select an action. In order to make use of this algorithm, only the previous action selection methods should be replaced by this method. The state space consists of three parameters that are shown in Fig. 3a: \ud835\udee5m Orthogonal distance of each agent from the bisector of the line sector connecting the other two agents. \ud835\udee5n Orthogonal distance of each agent from the line passing the other two agents. \ud835\udee5a The distance between the two other agents. \ud835\udefd is an angle made by a line passing two other agents, and the horizontal line and is used for simplification and symmetrization. Also, the agent is always in the first quadrant of the coordinate plane and in other cases, symmetrization is applied (Fig. 3b). For the action space, the allowable range is [90, 270] and divided into equal distances and the applied action constitutes the movement in one of these angles. The membership functions of fuzzy system inputs are shown in Fig. 4. The agents will receive a + 10 reward, if they reach the goal and create the formation, otherwise the reward will be \u22120.01. Reinforcement function = { +10 goal state \u22120.01 others (15) For simulation, agents are randomly placed in an environment of 20 \u00d7 20 dimensions and learning takes place as explained earlier"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003666_s00022-015-0274-2-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003666_s00022-015-0274-2-Figure9-1.png",
        "caption": "Figure 9 a A tetrahedron Pd = dd1d2d3 with one vertex d of the regular tetrahedron P = abcd, where the left side figure is the rotated Pd, b a subdivision of faces of Pd, c creases for a continuous flattening of Pd, d a folded state Pd by applying valley folds to {dw, d1pt, d2pt, d3pt} and a mountain fold to wpt and e a flat folded state of Pd",
        "texts": [
            "1, we may continuously flatten a polyhedron Q obtained by cutting off vertices of P so that at least two of its triangular faces are not necessarily equilateral triangles. Lemma 4.1. Let P = abcd be a regular tetrahedron and Pd = dd1d2d3 be a tetrahedron where d1, d2 and d3 are points on dc, db and da, respectively. Let w be the point on d1d2 so that dw divides \u2220d1dd2 into halves. Then Pd can be continuously flattened so that two faces dba and dca have no crease and so that the face dd1d2 has one crease on dw during the motion (see Fig. 9). Proof. Let Fi (1 \u2264 i \u2264 4) be faces of a tetrahedron W . Then it was proved in [9] that there are points vi in Fi (1 \u2264 i \u2264 4) such that each Fi is divided into three triangles by line segments from vi to vertices of Fi and such that each edge of W is a common edge of two congruent triangles that form a kite. If W is a regular tetrahedron, then those points vi are the centroids of Fi (1 \u2264 i \u2264 4). Let P = abcd be the regular tetrahedron defined in Lemma 2.3. Let Pd = dd1d2d3 be a tetrahedron and w be the point defined in Lemma 4.1 (see Fig. 9a). Let F1 = dd1d3, F2 = dd1d3, F3 = dd2d3, and F4 = d1d2 d3. It was shown in [9] that the line segments dvi (1 \u2264 i \u2264 3) divide the angle 1/3\u03c0 of the faces Fi at d into halves (see Fig. 9b). So, the point v1 is on the line segment dw. Apply Lemma 1 to the kite K = d1v1d2v4 in Fig. 9b. Then K has a flat folded state on the edge v1w, and there is a continuous motion by moving creases (edges) of {d1pt, d2pt} where pt is a point moving from w to v4 on the line segment wv4 according to t (0 \u2264 t \u2264 1). By applying valley folds to {dw, d1v4, d2v4, d3v4} the tetrahedron Pd can be continuously flattened, as shown in [9] (see Fig. 9c, d, f). Note that the continuous motion for the faces Fi (1 \u2264 i \u2264 3) in Lemma 4.1 is consistent with the continuous motion for P defined in Lemma 2.3. Theorem 4.2. Let P = abcd be the regular tetrahedron in Lemma 2.3. Let Q be a truncated tetrahedron of P with four hexagonal faces and four triangular faces, two of which are equilateral triangles. Assume that Q includes all the midpoints of the edges of P . Then Q is flattened by a continuous motion whose restriction to the four hexagonal faces of Q coincides with the continuous motion for P defined in Lemma 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001782_iros.2015.7354084-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001782_iros.2015.7354084-Figure5-1.png",
        "caption": "Figure 5. Walking support robot.",
        "texts": [
            " Torque-based Safety Device The features of the torque-based safety device are as follows: (i) If the torque of a shaft exceeds a preset threshold level, then the safety device for the shaft cuts off the torque transmission and switches off all of the robot\u2019s motors. We call the preset threshold level the \u201cdetection torque level\u201d. (ii) The detection torque level is adjustable. (iii) The safety device consists of only passive components without actuators, controllers, or batteries. With the above features, we expect the torque-based safety device to prevent the robot from exerting unexpected large forces on humans if the contact force-based safety devices do not work. Fig. 5 shows the walking support robot with velocity, torque, and contact force-based mechanical safety devices. The walking support robot has two drive units, two casters, a force sensor, and a contact force-based safety device. The force sensor is installed between the armrest and the body of Figure 2. Rescue of a human pressed against a wall by the robot. Human Wall Wall Walking Support Robot (Without Safety Devices) (With Safety Devices) Figure 1. Unexpected high speed robot motion. High Speed Walking Support Robot Human Lock Slope Slope (Without Safety Devices) (With Safety Devices) Figure 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002958_icma.2016.7558759-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002958_icma.2016.7558759-Figure2-1.png",
        "caption": "Fig. 2 Forbidden zone for thrusters",
        "texts": [
            " Take 5th and 6th thrusters in Fig. 1 for example, when the 5th thruster is located at the direction of the component of wake flow which is generated by the 6th thruster, the water flow will have a negative impact on the thrust force of 5th thruster. It leads to a substantial decline in thruster efficiency, and the maximum loss can reach forty percent. Therefore, the azimuth angle of thruster should not be allowed in that restricted area called forbidden zone to avoid loss of thrust as far as possible, as shown in Fig.2. In Fig.1, [- ,+ ] is the forbidden area for 5th and 6th thrusters. It must be pointed that the forbidden area is the azimuth angle of thruster, other than the angle of thrust force. Their quantitative difference is 180\u00b0. AFSA is a modern heuristic swarm intelligent algorithm which can be used to solve constrained optimization problem as thrust allocation. It has many advantages such as strong robustness, good global convergence and insensitivity to the initial value. As with other swarm intelligent algorithms AFSA works by a population of the so called artificial fish"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002220_6.2016-1529-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002220_6.2016-1529-Figure2-1.png",
        "caption": "Figure 2: Ornithopter used in this study (DelFly II) and definition of the body-fixed coordinate system (xb, yb, zb).",
        "texts": [
            " (1), and the cost function is now minimised with respect to the parameters, which are assumed to be unknown again. The minimisation of the cost yields the parameter update step. In this work, the minimisation process was performed using a Gauss-Newton algorithm. Initial guesses for the parameters were obtained using ordinary least squares regression. Further details on the application of the estimator are given in the subsequent sections, while more exhaustive information on ML estimation can be found in the literature.39\u201341 The subject of this work is the DelFly II, shown in Figure 2, a FWMAV developed at Delft University of Technology. The DelFly has four wings arranged in an \u2018X\u2019 configuration, that flap at frequencies between 10Hz and 14Hz, and an inverted \u2018T\u2019 tail for static stability and control purposes. It is characterised by an extensive flight envelope, enabling it to hover, but also to fly at velocities of up to 7m/s. The specimen used here has a wing span of approximately 274mm and a mass of approximately 18g. It can be controlled by means of variations in the flapping frequency, and deflections of the elevator and rudder surfaces in the tailplane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001029_adprl.2014.7010632-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001029_adprl.2014.7010632-Figure2-1.png",
        "caption": "Fig. 2. The body coordinates of the quadrotor where \u03c6, \u03b8, and \u03c8 are a roll, pitch, and yaw angle, respectively.",
        "texts": [
            " Then, we obtain as follows: QFL o (x\u0302k, v L k ) = [ z\u0302k vLk ]T [ HFL o (z\u0302z\u0302) HFL o (z\u0302vL) HFL o (vLz\u0302) HFL o (vLvL) ] [ z\u0302k vLk ] = [ z\u0302k vLk ]T HFL o [ z\u0302k vLk ] , (40) where HFL o (z\u0302z\u0302) =M\u0302T C\u0302TQoC\u0302M\u0302 + M\u0302T A\u0302T P\u0302LA\u0302M\u0302 , (41) HFL o (z\u0302vL) =M\u0302T A\u0302T P\u0302LB\u0302, (42) HFL o (vLz\u0302) =HT FL o (z\u0302vL), (43) HFL o (vLvL) =Ro + B\u0302T P\u0302LB\u0302. (44) Thus, we have the following control policy: vLk = \u2212H\u22121 FL o (vLvL) HT FL o (z\u0302vL)z\u0302k. (45) An algorithm of RL-based adaptive optimal output feedback control with L time delays is obtained by replacing line 4 of Algorithm 1 with vLk = \u2212F\u0302L o,iz\u0302k+e, and line 8 with F\u0302L o,i+1 = H\u0302\u22121 Fo,i(vLvL) H\u0302T Fo,i(z\u0302vL). We consider a quadrortor which has four rotors in cross configuration. Its lifting force is controlled by the speeds of the rotors. Shown in Fig. 2 is the body coordinate of the quadrotor. RL has been applied to the quadrotor for accommodating nonlinear disturbances [26], and to a helicopter, in order to design controllers for low speed aerobatic maneuvers [27]. However, the digital control of the quadrotor with taking the computation time into consideration has not been studied. In this paper, we apply the proposed control method to attitude control of the quadrotor at the hovering state and show its efficiency by simulation. Shown in Table I is a list of the quadrotor\u2019s system parameters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003365_phm.2016.7819880-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003365_phm.2016.7819880-Figure7-1.png",
        "caption": "Figure 7. Four types sun gear considered",
        "texts": [
            " An accelerometer (with sensitivity of 100mV/g and frequency range 0-10kHz) and a shaft encoder (produced by Encoder Products Co. with 1 pulse/revolution) are used for capturing vibration and Tacho signals simultaneously. The data are captured under speed up conditions from 0 RPM to 3000 RPM within 8 seconds data is used for the following analysis. The sampling frequency was set to be 7680Hz to accommodate all interested frequency contents of this test rig. The whole set-up arrangement is shown in Fig. 6 [6]. In the experimental process, four types of faulty sun gear are considered, see in Fig. 7. In the following, for demonstration purpose, vibration data from the tooth-missing fault (Fig. 7 (d)) is used to verify the effectiveness of the signal selection scheme. The physical parameters for planetary gearbox are listed in TABLE II in which gear teeth, number of planet gear and transmission ratio are given and calculated. In the experimental set-up, the ring gear of planetary gearbox is stationary and the sun gear is the input of the planetary gearbox system. Another very important calculation to the planetary gearbox is its characteristics orders. The characteristics orders are listed in TABLE III [7]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000900_s10846-014-0157-z-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000900_s10846-014-0157-z-Figure4-1.png",
        "caption": "Fig. 4 Direction of repulsive force",
        "texts": [
            " 32, rX4 is the radius of the vehicle (distance from the center to the tip of any propeller), Ladd is an additional distance to be more conservative in the repulsive distance. Thus, the repulsive force is frio = { Mri,o (I + \u03b2R) \u03be i\u2212\u03be o Lio Lio < Lrep 0 Lio \u2265 Lrep (35) where I \u2208 R 2 denotes the identity matrix, \u03b2 \u2208 (0, 1) is a constant and R \u2208 R 2 is a rotation matrix that redirects the repulsive force vector in a direction tangent to the obstacle R = [ 0 \u2212 sign(\u03b1) sign(\u03b1) 0 ] To better illustrate the proposed repulsive function (35), the magnitude of the repulsive force Mri,o (33) is shown in Fig. 3. The lines of the repulsive field are shown in Fig. 4, in this example \u03b2 = 0.5. A safer collision avoidance can be achieved by minimizing the distance toward the obstacle Lx once the vehicle enters in a repulsive zone. This can be accomplished by changing the nested controller constants k2, k3, k4 and b1 just at the moment of entering into a repulsive zone. These constants can be returned to their previous values when the vehicle abandons the repulsive zone. It is important to note that Lx in general is not obtained from x(t) in Eq. 31, but results from solving the system (19) with input (22), because it could consider any initial condition"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure2-1.png",
        "caption": "Fig. 2 Definition direction for tolerance model a toroidal direction and poloidal direction, b radial direction and VV structure",
        "texts": [
            " Another tolerance analysis about plates together with bolts if could meet whole block tolerance requirement had been finished in another report, so in this paper the block will be taken as acceptable product, and only the tolerance analysis about gaps between blocks and gaps between blocks with VV will be performed. The assembly procedure of local IWS model is that installing the under bracket on VV rib firstly, then matching the plate group and upper bracket, and installing them on VV rib through upper bracket. For purpose of analysis models could be described clearly, three coordinate directions were defined, as shown in Fig. 2. This analysis models is made in the global coordinate system, toroidal direction (X) is the latitude line direction of the VV inner shell, radial direction (Y) is the radius direction of the VV inner shell, and poloidal direction (Z) is the longitude direction of the VV inner shell [5, 6]. And consistent with each other, the toroidal gaps, poloidal gaps and radial gaps also have been confirmed as Fig. 2. After tolerance analysis models of an assembly are established in the complex solid design models. The reference feature (benchmark) of the assembly should be defined firstly; then correlative dimension chains of every parts among assembly were added orderly, and error limits of every parts were determined according to actual process capacity; constrained relationships between parts were applied; geometric dimensioning allowance of every parts had been finished; check the tolerance analysis results till they could met assembly requirements"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.11-1.png",
        "caption": "Fig. 4.11 Global and local systems of coordinates",
        "texts": [
            " The expression global coordinate system, which is represented in planar motion by two orthogonal axes that are rigidly connected at a point called origin of this system, is used to represent the inertial frame of reference. In the present work, the planar global coordinate system is denoted by xy. In addition, a body-fixed or local coordinate system is considered to define local properties of points that belong to a body i. This local system of coordinates is, in general, attached to the center of mass of the bodies and is denoted by \u03bei\u03b7i. This local system translates and rotates with the body motion, consequently, its location and rotation vary with time. Figure 4.11 shows the global and local systems of coordinates in a multibody system composed by nb rigid bodies. With the purpose of defining the geometric configuration of a multibody system, it is first necessary to select the type of coordinates to be used. In this work, due to their simplicity and easiness computational implementation, the \u201cabsolute\u201d or \u201cCartesian\u201d coordinates will be almost exclusively employed to formulate the equations of motion of multibody systems. Their relative advantages and drawbacks, often dependent on the application, were object of discussion, as stated earlier. However, it is worth noting that the absolute coordinates have the great merit to be quite straightforward, even for systems with high level of complexity. If a planar multibody system is made of nb rigid bodies such as the one illustrated in Fig. 4.11, then the number of absolute coordinates is n = 3 \u00d7 nb. Thus, the vector of generalized coordinates of this system can be written as q \u00bc qT1 qT2 . . . qTnb T \u00f04:22\u00de Let now consider a single body, denoted as body i, that is part of a multibody system, as Fig. 4.12 depicted. When absolute coordinates are used, the position and orientation of the body are defined by a set of translational and rotational coordinates. Thus, body i is uniquely located in the plane by specifying the global position, ri, of the body-fixed coordinate system origin, Oi, and the angle \u03d5i of rotation of this system of coordinates with respect to the x-axis of the global coordinate system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002253_ecce.2014.6953966-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002253_ecce.2014.6953966-Figure3-1.png",
        "caption": "Fig. 3. Geometry of the idealized machine and flux lines during the finite element analysis",
        "texts": [
            " the differential resistances, could be different from zero since they are affected by the mutual inductances; \u2022 the real part of z\u0307dq could be also negative, depending on the sign of the mutual inductances. In order to better investigate the behaviour of these parameters, an idealized example is considered in the next Section. At first, to understand better the equations derived above, it is instructive to consider a simple ideal machine and to compute its parameters by means of finite element analysis. Such machine is composed of four coils, two on the stator and two on the rotor, with the magnetic axis on the d- or the q-axis. Fig. 3(a) shows a sketch of the geometry: stator and rotor are separated by an airgap. Then, for each axis there is a coil on the stator and a coil on the rotor. The rotor is not moving. The main data of the model are reported in Table I. The coupling of the four coils can be properly described by [ z\u0307dd z\u0307dq z\u0307qd z\u0307qq ] = [ Rsd + j\u03c9Lsd j\u03c9MDQ j\u03c9MDQ Rsq + j\u03c9Lsq ] + \u03c92 (Rrd + j\u03c9Lrd)(Rrq + j\u03c9Lrq) + (\u03c9Mdq)2 \u00b7 \u00b7 ( (Rrd + j\u03c9Lrd) [ M2 Dq MDqMQq MDqMQq M2 Qq ] + (Rrq + j\u03c9Lrq) [ M2 Dd MDdMQd MDdMQd M2 Qd ] + \u2212 j\u03c9Mdq [ 2MDdMDq MQdMDq +MDdMQq MQdMDq +MDdMQq 2MQdMQq ]) (5) the d-q model derived above",
            " The q-axis high frequency parameters Lsq and Lrq decrease as the q-axis current magnitude increases due to the saturation on the q-axis. Fig. 4(b) shows the mutual inductances between each couple of coils. As expected the mutual inductance M between the stator and rotor coil on the same axis is almost constant because of the high saturation of the d-axis. As far as the mutual coupling between coils on different axes are concerned, their values are very close each other and lower than M . Then, it is worth noticing that they are both positive and negative, depending on the value of the q-axis current. Fig. 3(b) shows the flux lines due to the steady state currents Isd=10 A and Isq=-0.5 A. The magnetic field axis is highlighted by an arrow. The operating point has been selected so that the d-axis flux path results strongly saturated. This can be noted also observing the flux lines in Fig. 3 which cross the slots of the d-axis rotor winding even if no current is there imposed. Fig. 3(c) shows the flux lines due to a small-signal variation in the stator d-axis current only. In particular a pulsating field along the d-axis has been imposed. It can be noted that the flux lines are not only along the d-axis (as would be expected in an unsaturated machine with a d-axis current variation), but there is also a positive q-axis flux linkages. This is also reflected in the positive mutual inductances reported in Fig. 4(b), so that for a positive d-axis current a positive q-axis flux linkage is expected. In Fig. 3(c) the magnetic flux axis is also highlighted by an arrow superimposed to the flux lines. Then, the rotor ports have been short circuited and (8) have been used to compute the machine parameters as they result from the stator ports. The considered working point is again Isd=10 A and Isq=-0.5 A. The parameters are computed as: z\u0307dd = 2.011 + j 0.2304 z\u0307dq = \u22120.0089\u2212 j 0.0033 z\u0307qq = 3.782 + j 1.2711 z\u0307dq = \u22120.0089\u2212 j 0.0033 (9) It is worth noticing that the real part of z\u0307dq=z\u0307qd is negative. Such negative real parts in the machine parameters are not physical resistance but they represent a more complex interaction between the four coils of the machine at the high frequency"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.5-1.png",
        "caption": "Fig. 10.5 Design for assembly: (a) modularizing multiple parts into single sub-assemblies and (b) designing open enclosures for higher accessibility of important components",
        "texts": [
            " A product consisting of many different parts consumes large amounts of cost and time of production, and the chances for error in the assembly process is higher, too. In case there is one defective part assembled in the product, the cost of examination and disassembling the product may be even higher than producing a new product. Therefore, it is better to modify the product design to one with fewer components. Another strategy is to assemble the product in, say, two stages, with the earlier stage assembling only components into modules and the later stage 270 10 Design for Manufacturing assembling these modules into the final product as shown in Fig. 10.5a. Moreover, the important components should not be placed in a confined space because the higher accessibility of these components would enable replacement and maintenance if required in the future (Fig. 10.5b). Additionally, they should not be sealed with other parts of the product. If fasteners or screws are necessary, they should be located away from any obstructions with a wide opening region or on the product surface for higher accessibility. In fact, nowadays, there are products (especially high-technology consumer electronics) designed with very high accessibility to key components included due to a strategic marketing concern. Manufacturers wish to prevent their products from being copied illegally"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure2-1.png",
        "caption": "Fig 2 shows two kinds of air gap eccentricity: (a) static eccentricity; (b) dynamic eccentricity. For the static eccentricity, the rotor axis position is static relative to stator axis. For the dynamic eccentricity, the rotor axis is concentric with stator axis, but the rotating pivot point is off-center from rotor axis. Eccentric position is changing with rotating rotor.",
        "texts": [],
        "surrounding_texts": [
            "I. INTRODUCTION\nHigh speed permanent magnet machine is widely used in the FESS due to the advantages of high torque density and high efficiency. Comparing with the ordinary machine, external rotor machine is easier integrated with the flywheel, which makes the whole system compact and sturdy [1]. However, due to the magnetic levitation bearings and the heavy flywheel outside, the air-gap eccentricity phenomenon of the external rotor machine usually appears. Air-gap eccentricity can cause air gap magnetic field waveform distortion, affect the performance of the machine, such as magnetic saturation, back EMF asymmetric, additional loss, unbalanced magnetic pull, vibration noise and other issues. With the flywheel rotor operating in a vacuum, radiation becomes the primary heat transfer manner for removing losses incurred in the rotor [2]. In order to ensure the high speed machine run reliably, it becomes important to investigate additional rotor loss caused by the air gap eccentricity.\nThe problems related to air gap eccentricity have been studied for more than a century. All these studies were concerned with the unbalance magnetic pull (UMP), vibration noise and the methods to detect the eccentricity [3]. The most investigations of rotor losses caused by air gap eccentricity were carried out for induction machines [4, 5]. An extensive study on the effect of\n *Project supported by the National Natural Science Foundation of China\n(Grant No. 51307029 and No.51077023)\neccentricity on the rotor losses in permanent magnet machine is obviously missing from the literatures [6, 7].\nIn this paper, both an analytical method and finite element (FE) model are used for calculating the additional eddy current losses caused by air-gap eccentricity inside the rotor of high speed permanent magnet machines which is rotating at 20,000rpm with 100 kW rated power. The cross section of an external rotor machine with eccentric magnetic poles is shown in Fig 1. First, the analytical calculation of additional air-gap magnetic field and eddy current loss incurred in the rotor magnet under no-load conditions is derived, which is caused by the air-gap eccentricity. The relationship between additional eddy loss and eccentricity ratio is presented. Then the finite element simulations are used to analyze the rotor loss and the eddy current distribution respectively due to different types of eccentricity (static and dynamic eccentricity) and eccentric angle.\nII. ANALYTICAL CALCULATION\nA. Conventional air-gap magnetic field\nFor a surface mounted PMSM, the rotor MMF in the air-gap can be expressed by\n1 1 1\n1\n( , ) cos( )\ncos( )\nr r r\nr r\nf t F p t\nF p t\n\n  \n   \n   \n  \n   (1)\nWhere  is the harmonic order of rotor MMF, 2 1n   ,\n1 2n  , ...; rF  is the amplitude values of th rotor MMF; p is\nthe number of magnet pole pairs; 1 is the synchronous angular\nfrequency; r is the initial phase angle of th rotor MMF.\nThe stator MMF can be expressed by",
            "1 1 1\n1\n( , ) cos( )\ncos( ) s s s\ns s\nf t F p t\nF p t \n  \n   \n   \n  \n   (2)\nWhere  is the harmonic order of stator MMF, 6 1n   , 1 2n   , ...\nsF  is the amplitude values\nof th stator MMF. s is the initial phase angle of thv stator MMF. The magnetic permeance of air gap for non-eccentric rotor can be expressed by\n0 1( ) cos( )k k kZ      (3)\nWhere 1Z is the number of stator slot, 0 , k are the constant\ncomponent and thk component of magnetic permeance caused by stator slotting.\nThe flux density in the air gap without rotor eccentricity can be expressed by\n  1 1 1\n1 1 1\n0 1\n( , ) ( , ) ( , ) ( )\ncos( ) ( , )\ncos( ) ( , )\ncos( )\nr s\nr r r\ns s s\nk k\nb t f t f t\nF p t F t\nF p t F t\nkZ\n \n \n \n    \n   \n   \n  \n \n               \n     \n\n\n\n(4)\nNeglect the interaction components which are caused by the stator slotting and high harmonic order MMF on rotor and stator, the harmonic magnetic field can be expressed by\n1 1 1 1 0cos( ) cos( ) ( , )r r k s k F p t kZ F t            (5)\nWhere the first item is the harmonic magnetic field caused by sta tor slotting. It is the main harmonic which induces rotor eddy los ses in no-load. The second item is the space harmonic magnetic f ield caused by stator windings. It is the main harmonic which ind uces rotor eddy losses in load.\nB. Air-gap magnetic field for eccentric rotor\nThe equivalent air gap for eccentric rotor can be expressed by\n0( ) cos( )e et        (6)\nWhere 0 is the length of air gap for non-eccentric rotor; e is\nthe eccentric distance; e is the rotational angular velocity of\neccentric air gap, for the static eccentricity 0e  , for the\ndynamic eccentricity 1 e p\n    ;  is the rotational\nangular velocity of rotor. Considering stator slotting, the magnetic permeance of air gap for eccentric rotor can be expressed by\n0 1 0( ) cos( ) cos( )e k e k kZ t           (7)\nWhere  is the degree of eccentricity, which is defined as\n0\ne  \nAssume the MMFs of rotor and stator windings are not changing for eccentric rotor. The combination of equations (1),(2) and (7) leads to the air gap flux density for a machine with an eccentric rotor.\n \n1 1 1\n1 1 1\n0\n( , ) [ cos( ) ( , )\ncos( ) ( , )]\ncos( )\ne r r r\ns s s\ne\nb t F p t F t\nF p t F t\nt\n \n \n    \n   \n  \n    \n  \n \n\n (8)\nC. Permanent Magnet Eddy Current Loss\nDefined s is a slip, which can be expressed as\n1s           (9)\nWhere   is the rotational angular velocity of harmonic magnetic field. So the equation (8) can be divided into 4 parts: 1) Fundamental component of PM magnetic field\n\n\n1 1 1 1 0\n1 0 1 1\n1 1\n( , ) cos( ) cos( )\n1 cos[( 1) ( ) ] 2 cos[( 1) ( ) ]\ne r r e\nr e r\ne r\nb t F p t t\nF p t\np t\n      \n    \n   \n   \n    \n    \n(10)\nThe harmonic order is 1\n1 p  .The slip relative to rotor\nis 1 1 1 1 es p       ( ) , so the flux density relative to rotor is\n1 1_ 1 0 1 1\n1\n1 1 1\n1\n1 ( , ) cos[( 1) ( ) ]\n2 1\ncos[( 1) ( ) ] 1\ne rotor r e r\ne r\ns b t F p t\ns\ns p t\ns\n     \n   \n      \n      \n \n(11)\n2) Harmonic component of PM magnetic field",
            "2 1 0\n0 1\n1\n( , ) cos( ) cos( )\n1 {cos[( 1) ( ) ] 2\ncos[( 1) ( ) ]\ne r r e\nr e r\ne r\nb t F p t t\nF p t\np t\n  \n  \n\n       \n     \n    \n   \n    \n    \n\n (12)\nThe slip relative to rotor is 1 2 1 1 es p        ( ) , so the flux\ndensity relative to rotor is\n2 2_ 0 1\n2\n2 1\n2\n1 ( , ) {cos[( 1) + ( ) ]\n2 1\ncos[( 1) + ( ) } 1\ne rotor r e r\ne r\ns b t F p t\ns\ns p t\ns\n  \n\n      \n    \n    \n    \n\n(13) 3) Fundamental component of armature magnetic field\n3 1 1 1 0\n1 0 1 1\n1 1\n( , ) cos( ) cos( )\n1 {cos[( 1) ( ) ] 2 cos[( 1) ( ) ]}\ne s s e\ns e s\ne s\nb t F p t t\nF p t\np t\n      \n    \n   \n   \n    \n    \n(14)\nThe harmonic order is 1\n1 p  .The slip relative to rotor\nis 1 3 1 1 es p       ( ) , so the flux density relative to rotor is\n3 3 _ 1 0 1 1\n3\n3 1 1\n3\n1 ( , ) {cos[( 1) ( ) ]\n2 1\ncos[( 1) ( ) ] 1\ne rotor s e s\ne s\ns b t F p t\ns\ns p t\ns\n     \n   \n     \n     \n(15)\n4) Harmonic component of armature magnetic field\n4 1 0\n0 1\n1\n( , ) cos( ) cos( )\n1 {cos[( 1) ( ) ] 2\ncos[( 1) ( ) ]\ne s s e\ns e s\ne s\nb t F p t t\nF p t\np t\n  \n  \n\n       \n     \n    \n   \n    \n    \n\n (16)\nThe slip relative to rotor is 1 4 1 1 es p        ( ) , so the flux\ndensity relative to rotor is\n4 4_ 0 1\n4\n4 1\n4\n1 ( , ) {cos[( 1) + ( ) ]\n2 1\ncos[( 1) + ( ) ]} 1\ne rotor s e s\ne s\ns b t F p t\ns\ns p t\ns\n  \n\n      \n    \n    \n    \n\n(17) The combination of equations (11), (13), (15) and (17) leads to the flux density relative to rotor for a machine with an eccentric rotor.\n_ 1_ 2_\n3_ 4_\ne rotor e rotor e rotor\ne rotor e rotor\nb t b t b t\nb t b t\n  \n \n \n \n( , ) ( , ) ( , )\n( , ) ( , ) (18)\nCalculate the additional eddy current losses in permanent magnet which is caused by eccentric rotor. The analytical modeling showed in Fig3 is based on the following assumptions: a) Neglect end effect; b) Neglect eddy current reaction; c) The magnetic field in air gap only has radial component.\nBased on polar coordinate 2-D model, the magnetic vector potential in permanent magnet can be given as\n_e rotorA r Bd r b t d     ( , ) (19)\nThe eddy current in the magnet can be calculated by A\nJ t\n   \n (20)\nWhere is the conductivity of magnet. The combination of equations (19) and (20) leads to the eddy\ncurrent density in magnets.\n1 1_ 2 2_\n3 3_ 4 4_\ne rotor e rotor\ne rotor e rotor\nJ r s b t s b t\ns b t s b t\n  \n \n  \n \n[ ( , ) ( , )\n( , ) ( , )] (21)\nThe magnet loss caused by eccentric rotor can be calculated by\n2 2 2 3 3 0\n2 2 2 2 2 2 2 2 1 1 2 3 1 4\n1\n1\n24\ne eddy\nb a\nr r s s\np L J J dS\nL R R\ns F s F s F s F\n\n   \n\n   \n \n \n   \n\n \n*\n_\n( )\n( )\n(22)\nWhere L is the active length of magnet.  is the width of\nmagnet. aR , bR are the inner radius and external radius of magnet, respectively. For the different kinds of air gap eccentricities, the components in harmonic magnetic field are different. For the\nstatic eccentricity 0e  , 1 3\n1\n1 s s p\n  \n , 2\n1\n1 s p    ,\n4 1 1\np s\np    . For the dynamic eccentricity e  ,\n1 2 3 0s s s   , 4 1\n1\np s\np    .\nBased on the analysis above, it comes to a conclusion that both rotor magnetic field and armature magnetic field can induce eddy current loss in magnet under static eccentricity. For the dynamic eccentricity, only armature harmonic magnetic field will induce"
        ]
    },
    {
        "image_filename": "designv11_64_0002410_0309524x16647842-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002410_0309524x16647842-Figure3-1.png",
        "caption": "Figure 3. Accelerometer mounting locations on the rear PMG housing.",
        "texts": [
            " The general arrangement of main equipment is illustrated in Figure 2. Mechanical and electrical imbalances can be studied on the test rig. To mimic a rotor imbalance, a mass block was bolted to the load disc. To create electrical imbalance, the three-phase resistors can be varied. Piezoelectric accelerometers were used due to their wide linear frequency range, broad dynamic amplitude range, and suitability for limited space inside the PMG. The four locations of the accelerometers are shown in Figure 3. Location 1 is at 3 o\u2019clock outside of the bearing housing. Location 2 is at 3 o\u2019clock outside of the PMG cover. Location 3 is at 6 o\u2019clock outside of the bearing housing. Location 4 is at 6 o\u2019clock outside of the PMG cover. For a shaft speed of 9 Hz, the acceleration spectra for the four different mounting locations are shown in Figure 4. Clearly, Location 3 below the bearing and within the load zone gives the most useful data for condition monitoring. The signal power is the largest at this location, which makes intuitive sense because this location is inside the load zone",
            " The lower frequency matches with fr . Energy at frequency sub-band 9 Hz increased with the electrical imbalance. However, mechanical and electrical imbalances cannot be distinguished by this method. Voltage signals from the PMG were collected simultaneously with the vibration signal. FFT of voltage was analyzed for the mechanical and electrical imbalances. For the mass imbalance used in Figure 8, the FFT of voltage signals collected for fr = 9Hz is shown in Figure 10. All significant peaks occurred at the same frequencies as in Figure 3. The voltage magnitude at 74.75 Hz, which is the power line frequency, changed from 3.991 to 4.121 V which is about a 0.75% difference. This means that mechanical imbalance does not show obvious changes in the magnitude of the voltage signals. An electrical imbalance was simulated as for Figure 9 by reducing one-phase resistance from 3.6 to 1.8 \u03a9 and then to 0.9 \u03a9. The resistance on the other two phases remained at 3.6 \u03a9. The FFT of voltage signals collected at PMG running at 9 Hz was computed and the results are shown in Figure 11. The significant peak values shown in Figure 11 are different from the ones in Figure 3. Results are compared in Table 1. The voltage magnitude at fr decreased when the imbalanced phase resistance was reduced. The voltage magnitude at 3 fr increased with a reduction in resistance which means electrical imbalance increased. Further analysis is required to investigate the changes at fr and 3 fr with electrical imbalance. The electrical imbalances were seen from the voltage analysis, but the changes are small compared to the changes in the load resistance. Results for variable shaft speed Under a sinusoidal speed variation of fr with mean value 9 Hz, magnitude 1 Hz, and frequency 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.8-1.png",
        "caption": "Fig. 3.8 Alternative configuration for the derivation of the kinematics relation. Note that the deformation is exaggerated for better illustration",
        "texts": [
            "7 Segment of a beam under pure bending in the x\u2013z plane. Note that the deformation is exaggerated for better illustration If these relations for the circular arcs are used in Eq. (3.12), the following results: \u03b5x = (R + z)d\u03d5y \u2212 Rd\u03d5y dx = z d\u03d5y dx . (3.15) From Eq. (3.13) d\u03d5y dx = 1 R results and together with relation (3.11) the strain can finally be expressed as follows: \u03b5x(x, y) = z 1 R (3.11)= \u2212z d2uz(x) dx2 (3.11)= z\u03ba. (3.16) An alternative derivation of the kinematics relation results from consideration of Fig. 3.8. From the relationof the right-angled triangle 0\u20321\u20322\u2032, thismeans11 sin\u03d5y = ux z , the following relation results for small angles (sin\u03d5y \u2248 \u03d5y): ux = +z\u03d5y. (3.17) Furthermore, it holds that the rotation angle of the slope equals the center line for small angles: tan\u03d5y = \u2212 duz(x) dx \u2248 \u03d5y. (3.18) If Eqs. (3.18) and (3.17) are combined, the following results: ux = \u2212z duz(x) dx . (3.19) The last relation equals (ds \u2212dx) in Eq. (3.12) and differentiation with respect to the x-coordinate leads directly to Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure7.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure7.8-1.png",
        "caption": "Fig. 7.8 Printed object (black) and supporting material (white) during fabrication. (Reproduced with permission from Ultimaker)",
        "texts": [
            " CAD software can then transform the geometry into multiple slices of two-dimensional cross sections (Fig. 7.7). During this step, designers can configure the slicing procedures with the machine parameters such as the layer thickness, printing position, object position, and supporting mode. Designers should also select the printing materials. If necessary, support materials 198 7 Medical Imaging and Reverse Engineering can be added underneath the product shape as temporary supports during the fabrication process (Fig. 7.8). After printing the material as defined by the STL model (and the support material), post-processing should then be conducted, including taking the object out of the printing board, removing the support material, cleaning the printed object, further curing (e.g., by additional ultraviolet exposure for photo-curing materials), surface coating, assembling, and final finishing such as polishing over the object surface. Rapid prototyping fabricates some components with their geometry defined by computer-readable formats"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003395_icarcv.2016.7838849-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003395_icarcv.2016.7838849-Figure3-1.png",
        "caption": "Figure 3. Experimental setup - side view",
        "texts": [
            "2 and Table 1 is configured for one axis reaction wheel, it is also possible to customize it in order to get multiaxis reaction wheel systems. The main constants and parameters of experimental setup are listed in Table II. Fig.2 illustrates the one dimensional stabilization prototype that is built in order to assess the feasibility and develop the control algorithms for one dimensional indirect stabilization problem. The one-dimensional indirect stabilization system on which the experiments are conducted is illustrated in Fig. 3. This system has two actuators, which are flat type brushless DC motor and brush DC motor and shaft of the brush DC Motor is equipped with an optical encoder for calculating the linear displacement although it is not mandatory for braking operation. In addition, this setup has a two-axis MEMS rate sensor mounted on platform body to sense the disturbance about the line-of-sight. There is also an absolute encoder mounted on the revolute joint of the platform. Since the stabilization controller is designed for steady pointing of during operation, the base of the platform is actuated to induce the required disturbance at the base of the platform, in a way similar to real-life terrain navigation applications"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003273_vppc.2016.7791798-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003273_vppc.2016.7791798-Figure2-1.png",
        "caption": "Fig. 2. Structure of Cooperative Driving Control System for Vehicles",
        "texts": [
            " Hence, a simple 2DOF kinematic model is used in plan level while a 3DOF dynamic model is adopted in tracking control level. In order to reduce the computational burden and control complexity, vehicles in cooperative driving system are assumed homogeneous, i.e. the parameters and response of each vehicle are identical. Hence, the multi-vehicle control system can be translated to single vehicle control problem, if the distributed control is employed. The 2DOF kinematic model of vehicle which conforms to Ackerman steering law is shown in Fig. 2. It is assumed that the vehicle with a rigid body and no deformable wheels is the front wheel steering structure. The model can be represented as: \u23a7\u23a8 \u23a9 ?\u0307? = \ud835\udc63 cos(\ud835\udf13) ?\u0307? = \ud835\udc63 sin(\ud835\udf13) ?\u0307? = ?\u0307? (1) where (\ud835\udc65, \ud835\udc66) are the global coordinates of the rear wheels, \ud835\udc63 is the longitudinal velocity at the middle of the rear axis, \ud835\udf13 is the heading angle, \ud835\udc36\ud835\udc40 is the vehicle barycenter, and \ud835\udf14 is the yaw rate. As the flocking algorithm in [8] is based on a particle-linear model, it is necessary to linearize the nonlinear system (1) before using this algorithm to construct plan level controller"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003137_0954410016676847-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003137_0954410016676847-Figure2-1.png",
        "caption": "Figure 2. Lambert\u2019s problem in a two-dimensional reference frame in the orbital plane.",
        "texts": [
            " As illustrated in Figure 1, the angles and i can be determined by the unit vector of angular momentum 1h and the unit vector of Z-axis 1Z, that is, cos\u00f0 \u00de \u00bc 1X\u00f01Z 1h\u00de \u00f01\u00de cos\u00f0i\u00de \u00bc 1h1Z \u00f02\u00de where 1X\u00bc [1, 0, 0], 1Z\u00bc [0, 0, 1] and 1h \u00bc r1 r2= k r1 r2 k. at UNIV OF WESTERN ONTARIO on November 25, 2016pig.sagepub.comDownloaded from Once and i are found, the remaining four elements can be determined within the orbital plane. For convenience of description, a two-dimensional reference frame in the orbital plane is introduced. As shown in Figure 2, the x-axis is along the node line, which is the intersection of the orbital plane and the equatorial plane, and the y-axis is perpendicular to the x-axis and inside the orbital plane. Let c be the chord length between P1 and P2 and be the transfer angle. The initial position vector r1 and target position vector r2 can be explicitly expressed as follows: r1 \u00bc x11x \u00fe y11y \u00f03\u00de r2 \u00bc x21x \u00fe y21y \u00f04\u00de where 1x and 1y are the unit vectors of the x- and y-axes respectively, and x1, y1, x2, and y2 are the coordinates of P1 and P2 in the xy-system, which are known for Lambert\u2019s problem",
            " Lambert\u2019s problem based on semi-major axis a Elliptic orbits (0< e< 1) For elliptic orbits, when the semi-major axis a is used as the independent variable for Lambert\u2019s problem, the eccentricity e can be expressed in terms of a by solving equation (23) as follows e \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi C1 C2 p ffiffiffi 2 p ac \u00f025\u00de where C1 and C2 are defined as C1 \u00bc 2a2c2 a r1 \u00fe r2\u00f0 \u00de c2 r1 r2\u00f0 \u00de 2 \u00fe 1 4 r1 \u00fe r2\u00f0 \u00de 2 c2 c2 r1 r2\u00f0 \u00de 2 \u00f026\u00de C2 \u00bc 1 2 c2 r1 r2\u00f0 \u00de 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0r1 \u00fe r2\u00de 2 c2 q ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f02a s\u00de\u00bd2a \u00f0s c\u00de p \u00f027\u00de where the conditions r1\u00fe r2> c and r1 r2k k5 c have been used in deriving equation (27), and s is the semi-perimeter of the triangle OP1P2 in Figure 2, which is expressed as s \u00bc \u00f0r1 \u00fe r2 \u00fe c\u00de=2 \u00f028\u00de at UNIV OF WESTERN ONTARIO on November 25, 2016pig.sagepub.comDownloaded from For elliptic orbits, the minimum-energy orbit is an ellipse that passes P1 and P2 with the smallest possible value of the semi-major axis am. The semi-major axis of the minimum-energy orbit is am\u00bc s/2. For all elliptic orbits having a semi-major axis a and passing P1 and P2, we have a5 am. Using these facts, we deduce that 2a5 s, so equation (27) can be rewritten as C2 \u00bc 1 2 c2 r1 r2\u00f0 \u00de 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0r1 \u00fe r2\u00de 2 c2 q ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f02a s\u00de p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2a \u00f0s c\u00de p \u00f029\u00de Together with equations (13) to (16), equations (19) and (20), and equation (25), the transfer-time equation (10) can be rewritten as t \u00bc s 2a 3 2 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f02a s\u00des p 2a \u00fe 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f02a s\u00des p 2a \" \u00fe cot 1 ffiffiffiffiffiffiffiffiffiffiffiffiffi 2a s p ffiffi s p cot 1 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2a s p ffiffi s p \u00f030\u00de where the parameter is defined as \u00bc s c s \u00f031\u00de The parameter decreases monotonically from 1 to 0 as the transfer angle increases from 0 to , and increases monotonically from 0 to 1 as increases from to 2 ",
            ", is singlevalued, continuous, and monotonic for all conic orbits where the various values of the parameter l are applied here to represent different transfer angles. Assuming that the initial position vector r1 is fixed and r2\u00bc r1 where is a specified constant ratio, once l is given, the transfer angle can be obtained as follows \u00bc 2 arccos lsffiffiffiffiffiffiffiffi r1r2 p \u00f085\u00de thus the target position can be calculated by the transfer angle and the transfer-time function can be completely determined for a specific l. Finally, based on Figure 2, the true anomalies f1 and f2 can be represented as linear functions of ! as f1 \u00bc tan 1\u00f0 y1=x1\u00de ! \u00f086\u00de f2 \u00bc tan 1\u00f0 y2=x2\u00de ! \u00f087\u00de t \u00bc ffiffiffi 1 s \u00f0xc cos\u00f0!\u00de \u00fe yc sin\u00f0!\u00de\u00de\u00bdr1\u00f0x2 cos\u00f0!\u00de \u00fe y2 sin\u00f0!\u00de\u00de r2\u00f0x1 cos\u00f0!\u00de \u00fe y1 sin\u00f0!\u00de\u00de \u00f0r1 r2\u00de 2 \u00f0xc cos\u00f0!\u00de \u00fe yc sin\u00f0!\u00de\u00de 2 3 2 \u00f0r1 r2\u00de\u00f0 yc cos\u00f0!\u00de xc sin\u00f0!\u00de\u00de ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0r1 r2\u00de 2 \u00f0xc cos\u00f0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000971_s11012-014-0003-1-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000971_s11012-014-0003-1-Figure1-1.png",
        "caption": "Fig. 1 Model of the rotor system with lumped masses",
        "texts": [
            "02 0.025 0.03 0.035 -2.5 -2 -1.5 -1 -0.5 0 0.5 1 1.5 2 2.5 Friction coefficient D is pl ac em en t i n y di re ct io n (1 0 -3 m ) A A' B B' C C' D D' (b) Flexible rotor system running at 100Hz Fig. 3 Bifurcation diagrams of the transverse vibration amplitudes along the friction coefficients of the oil-block, a rigid rotor system running at 40 Hz and b flexible rotor system running at 100 Hz The schematic diagram of the Jeffcott rotor system with an oil-block inside the rotating drum is shown in Fig. 1. The rotor part, which is composed of a shaft and a hollow polymer drum, is supported by two rolling bearings and driven by a speedadjustable electric motor. The hollow polymer drum is installed on the middle point of the shaft. A volume of lubricant oil is injected in it. During the rotating process, the lubrication oil may form an oil-block which will rotate on the inner wall of the drum. During modeling the rotor system with oil-block inside the rotating drum, two assumptions about the oil-block are given as follows. (1) The oil-block is small enough to neglect its geometric shape influence during its revolution on the inner drum wall, once it forms a congregated block, (2) self-rotation of the oilblock is not considered. As shown in Fig. 1, the global coordinate system is set as Oxyz, and the z-axis is along the centerline of the two bearings. The transverse displacements of the lumped mass (namely the rotating drum and the equivalent shaft mass) at the drum center are (x, y) in x- and y-directions, and the corresponding angle displacement components are hx and hy. A moving coordinate system can be attached to the rotating drum. External excitation of the rotor system includes two parts: one is the unbalance force of eccentric rotor mass and the other is the interaction force induced by the oil-block",
            " It is seen obviously that when the mass is much smaller, NSV of the rotor system cannot be aroused. When the mass exceeds a threshold value of 10 9 10-3 kg, the NSV appears and its amplitude sharply increases, which is much lager than that of 19 rotating frequency. Moreover, the NSV frequency component is greater than that of the fundamental rotating frequency for the rigid rotor system. The values of nonsynchronous amplitudes for the both cases increase gradually and slightly as the mass of the oil-block increasing (Fig. 18). The measured vibrations of the corresponding test rig shown in Fig. 1 are described here. In the experiment, 2 ml water, 2 and 4 ml lubricant oil No. 32# are added into the drum, respectively. And the vertical and horizontal displacements of the shaft near the drum are measured with four non-contact eddy-current sensors. The frequency spectra of the measured transverse shaft displacements in the vertical direction are illustrated in Fig. 19 for the three cases. Compared with the case of 2 ml oil No. 32# in Fig. 19b, the NSVs are more obvious for the case of 2 ml water, and the amplitudes of which are much larger as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002199_978-0-262-33027-5-ch105-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002199_978-0-262-33027-5-ch105-Figure2-1.png",
        "caption": "Figure 2: PhysX\u2019s soft-body simulation is implemented using a tetrahedral mesh (a) which can be used to drive an accompanying triangular mesh for rendering (b).",
        "texts": [
            " The fundamental soft-body system used is available as a part of NVIDIA PhysX, but not in most versions, and (at the time of writing) not in recent versions. (The work presented here was implemented with SDK version 2.8.1 and PhysX System Software 9.11.0621.) This system offers well-established abilities to simulate rigid bodies and joints, and can do so in combination with the soft-body simulations which are the focus of this work. In this system, soft bodies are simulated as tetrahedral meshes (Figure 2a), with vertices simulated as particles, and additional constraints applied per tetrahedron to produce some of the more complicated effects described in the following subsections. For rendering, the simulated tetrahedral mesh can be used to deform a corresponding triangle mesh of arbitrary complexity (Figure 2b). Meaningful Morphological Complexity. Proceedings of the European Conference on Artificial Life 2015, pp. 604-611 One desired characteristic of soft body muscles is the preservation of volume, which produces a familiar squash-andstretch behavior as muscles decrease and increase in length. In PhysX, this property (which they call volume stiffness) is directly available as an attribute of simulated soft bodies, and is implemented as restorative movements applied to tetrahedral vertices (Figure 3a)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.19-1.png",
        "caption": "Fig. 3.19 Double wishbone wheel suspension with disconnected joints",
        "texts": [
            "; 6nB: \u00f03:10\u00de The advantage of this approach lies in the relatively simple formulation of the corresponding constraint equations, as well as in the weak-coupled structure of the whole system of constraint equations (only a few variables per equation, collected into a matrix with many zeroes, the so-called Sparse matrices). A disadvantage is the fact that the formulation of the function w\u00f0q\u00de is profoundly implicit. On the one side, this has a high numerical complexity for the solution of the system of Eq. (3.10); on the other side it can also frequently give numerical singularities during the numerical simulation, which are not necessarily caused by mechanically interpretable singularities (singular positions). Example (Fig. 3.19). Double wishbone wheel suspension The assembly yields two vectorial constraint equations for each joint: ri \u00fe irG bi\u00f0 \u00de \u00bc rj \u00fe jrG bj : \u00f03:11\u00de The total balance for the degrees of freedom is given as follows: 5 6 Equations of motion 30 \u00f03 5\u00fe 4 3\u00de Constraints -27 DoF 3 (one isolated DoF (Fig. 3.19)) Principle: disconnection of only one joint from each kinematic loop; transformation of a system with kinematic loops into a system with tree structure. Constraints: closure conditions for the kinematic loops. Here, the multibody system is transformed by the disconnection of suitable joints into a system with a tree structure. The kinematic description variables are introduced as nb \u00bc XnG i\u00bc1 fGi \u00f03:12\u00de auxiliary variables bj, which describe the relative position of two adjacent bodies. The dependencies between the auxiliary variables are implicitly stated by formulation of the closure conditions only at the disconnected joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001030_haptics.2014.6775481-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001030_haptics.2014.6775481-Figure4-1.png",
        "caption": "Figure 4: An actuator stretched the rubber sheet to change the force response",
        "texts": [],
        "surrounding_texts": [
            "As described above, it is possible to control the transition displacement between the force phases by stretching; however, the reaction force also changes depending on the stretching intensity. In order to render the desired force, a haptic display device generates an additional force and enhances the reaction force of the base object. Based on Eq. ??, the force that the haptic device generates Fdevice was determined by the difference between the model force of the target object Mtarget and the model force of the base object Mbase. In this study, Mtarget and Mbase were characterized as following: Mbase = abase 1 x2 +abase 2 x (11) Mtarget = atarget 1 x2 +atarget 2 x (12) where x is the pushing displacement, and a1,a2 are the coefficients. Fig. ?? shows the model force calculated by Eq. ??, and the measured static force response of the soft rubber sheet. The parameters used in the calculation are shown in Table ??. The force that the haptic device generates Fdevice can also be determined by Eq. ??: Fdevice = (atarget 1 \u2212abase 1 )x2 +(atarget 2 \u2212abase 2 )x. (13)"
        ]
    },
    {
        "image_filename": "designv11_64_0001284_j.apor.2014.07.012-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001284_j.apor.2014.07.012-Figure1-1.png",
        "caption": "Fig. 1. Definition sk",
        "texts": [
            " However, in literature, many different forms xist, and in particular those developed by MMG [24] have received wide acceptance because of their \u201cmore physical\u201d approach to he maneuverability problem. Thus, there is a need to study fuzzy ig. 2. Control over a straight line path (U = 15 knots, = 2, 0 = 0, simulation ime = 2500 s). control in conjunction with other maneuvering models, e.g. MMG models. 3. Target Path Iteration algorithm (TPIA) The \u2018Target Path Iteration\u2019 (TPI) algorithm for automatic control that is proposed in the present work is essentially the inverse problem of maneuvering in a limited sense. Consider Fig. 1 for a definition sketch. The ship parameters (as given in Tables 1 and 2) and its initial conditions (u, v, r and \u0131 at time t = 0) are known. The task of the algorithm is to find a new rudder angle \u0131 (or alternatively \u0131) at t = t such that the ship moves to a desired (or target) position given by x (or alternatively x) and y (or alternatively y). The essential steps of the algorithm to achieve this task are given below in a step-by-step manner. pplied F ( 1 \u221a\u221a q\u2211 S.K. Bhattacharyya, D.K. Gupta / A (i) Let the target path be given as a set of successive n points, whose coordinates in the earth-fixed horizontal (x, y) plane are given by (xi, yi, i = 1, 2, "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000395_978-94-007-4132-4_8-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000395_978-94-007-4132-4_8-Figure10-1.png",
        "caption": "Fig. 10. Hoecken\u00b4s sketch of Redtenbacher\u00b4s gear train",
        "texts": [
            " It is not necessary to write about the significance of Ferdinand Jakob Redtenbacher (1809- 1863) in the evolution of mechanical engineering to a science in Germany, because there is detailed and excellent description of his lifework, cf. [11, 12]. Hoecken explains that three of the seven gearwheels are not necessary for the correct function of the mechanism, but only serve to have the input rotary axis with the driving crank coincident with the output rotary axis (\u201creturning\u201d mechanism). Both axes have metal pointers in form of arrows for demonstration, similar to those of a clock. The base mechanism is that of a crank-rocker ABCD (Fig. 10a) with four two-bytwo mating gearwheels with their centres in B, C and D (Fig. 10b). The gearwheel centred in B forms the driving crank and rotates around A (eccentric wheel). Dependent on the dimensions of the crank-rocker and the radius r1 to r4 of the gearwheels it is possible to vary the output angle \u03c9 of the gearwheel 4 related to the input angle \u03b1 of the crank or gearwheel 1, i.e. the transmission ratio can be partly positive, zero (standstill of gearwheel 4) or negative (\u201cpilgrim\u00b4s step\u201d motion of gearwheel 4), while the crank 1 rotates continuously with constant angular velocity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.11-1.png",
        "caption": "FIGURE 8.11",
        "texts": [
            " A ball bearing has the low friction properties of rolling contacts and the high load capacity of revolute joints. Ball or roller bearings are kinematically equivalent to simple revolute joints. Another example is the universal joint shown in Figure 4.9(f). This is a combination of revolute joints and has two rotational degrees of freedom. Planar joints are much simpler to define mathematically than their counterparts in space. For example, a planar revolute joint allows relative rotation at a point P that is common to Bodies i and j, as shown in Figure 8.11(a). If one body is held fixed, the other body has only a single rotational DOF. Thus, a planar resolute joint eliminates two DOF from the pair. This joint is defined by locating point Planar joints. (a) J1: revolute/hinge/pin joint, (b) J1: prismatic/slider, and (c) J2: pin-in-slot. Mathematical formulation of a revolute joint: (a) planar revolute joint, (b) spatial revolute joint, and (c) dot-1 constraint. Higher pair joint: a cam-follower in the mechanism of an engine inlet or outlet valve. Pi on Body i by s0Pi in the x0i y0i frame (fixed to body i) and locating Pj on Body j by s0Pj in the x0j y0j frame (fixed to body j), respectively",
            "88) where Ai and Aj are the transformation matrices that transform position vectors s0Pi and s0Pj from their respective frames x0i y0i and x0j y0j to the global frame X-Y, respectively. Note that there are two equations in Eq. 8.88; therefore, motion is constrained in both the X- and Y-directions. Also note that the constraint equations of Eq. 8.78 in Example 8.8 define a revolute joint at points P1 and P2 of the crank and rod, respectively. A spatial revolute joint between Bodies i and j allows relative rotation about a common axis, but precludes relative translation along this axis, as shown in Figure 8.11(b). To define the revolute joint, the joint center is located on Bodies i and j by points Pi and Pj. The axis of relative rotation is defined in Bodies i and j by points Qi and Qj and hence unit vectors hi and hj along the respective z00-axes of the joint reference frames. The remaining joint definition frame axes are defined at the convenience of the designer. The analytical formulation of the revolute joint is that points Pi and Pj coincide and that body-fixed vectors hi and hj are parallel, leading to the constraint equations Fs Pi;Pj \u00bc ri \u00fe Ais 0P i rj Ajs 0P j \u00bc 0 (8.89a) Fp hi; hj \u00bc \" Fd f i; hj Fd gi; hj # \u00bc 0 (8.89b) There are three scalar equations in Eq. 8.89a, which eliminate three relative translational DOF between Bodies i and j at points Pi and Pj. Eq. 8.89b defines a parallel constraint, consisting of two so-called dot-1 constraints. A dot-1 constraint is defined by a dot product of two perpendicular vectors. In this case, fi is perpendicular to hj and gi is perpendicular to hj. Also, as shown in Figure 8.11(c), hi is perpendicular to both fi and gi; therefore, hi and hj are in parallel. There are, overall, five constraint equations in Eqs 8.89a and 8.89b; all are independent and allow only rotation along hi or hjdthus, a revolute joint in space. The DOF that each spatial joint eliminates are summarized in Table 4.5. Consider mechanical systems that are made up of a collection of rigid bodies in a plane, with kinematic constraints between them. The variational approach commonly found in dynamics textbooks is employed to formulate differential equations of motion",
            " The key idea is to couple the differential equations of motion with the kinematic constraint equations by introducing Lagrange multipliers to account for the constraints. The variational equation of motion for a rigid body in a plane can be formulated as drT m\u20acr F \u00fe dqT J0\u20acq n \u00bc 0 (8.90) where r is the position vector to the mass center of the body. F and n are external forces and torque, respectively. J0 is the polar moment of inertia at the centroid of the body referring to the body-fixed reference frame x0-y0 (see Figure 8.11(a)). dr and dq are the virtual displacements and rotation of the rigid body, respectively. The body-fixed reference frame x0-y0 is defined at the centroid of the body. Equation 8.90 can be written as dqT M\u20acq Q \u00bc 0 (8.91) where q \u00bc [r, q]T is the vector of generalized coordinates. Q \u00bc [F, n]T is the vector of generalized forces. M is the diagonal mass matrix consisting of mass m and moment of inertia J0 for the rigid body. The variational equations of motion for each Body i in a planar multibody system of nb bodies, given by Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001232_s1068798x15010190-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001232_s1068798x15010190-Figure2-1.png",
        "caption": "Fig. 2. Position of the contact spot and instantaneous con tact ellipse in localized contact.",
        "texts": [
            " In engagement with no edge contact, the results of the two solutions are practically the same. Hence, localization of contact is required. Contact is said to be localized if the instantaneous contact areas do not pass beyond the tooth contour even under the maximum load and with permissible error in the tooth pitch fpt. Lateral tooth surfaces are said to be optimal if they minimize the maximum (over all engagement phases) contact pressure \u03c3co in localized contact. The degree of contact localization is determined by three variables (Fig. 2): the distance \u0394a of the contact pattern from the upper longitudinal edge of the tooth; and the distances \u0394e and \u0394i of the contact pattern from its external and internal transverse edges, respectively. The distance from the edge is assumed to be nega tive if any contact pattern moves beyond the tooth contour and positive otherwise. The required distances \u0394ar, \u0394er, and \u0394ir are the ini tial data in optimization. For gears of precision classes 5 and 6, we assume that \u0394ar \u2265 mn/10, \u0394er \u2265 b/10, and \u0394ir \u2248 0",
            " These parameters provide the following information: the position of the center of the future contact pattern; the direction of the contact path on the tooth surface; the length of the major semiaxis of the instantaneous contact ellipse; and the nonunifor \u0394a1 \u0394a2\u2013 0.2Y;< Y \u0394a1 \u0394a2+ 1.8\u0394ar;>= \u0394i \u0394ir\u2013 0.3\u0394ir;< \u0394e \u0394er\u2013 0.2\u0394er;< f 0.7w0.< \u0394a1, \u0394a2, \u0394a1 \u0394a2+ 2.2\u0394ar.< RUSSIAN ENGINEERING RESEARCH Vol. 35 No. 1 2015 SYNTHESIS OF CONTACT OPTIMAL SPIRAL BEVEL GEARS 53 mity of transmission. The parameters employed in the synthesis are as follows. 1. The displacements \u0394\u03be and \u0394\u03b7 of the calculation point P on the lateral tooth surface relative to the mid point M of the tooth (Fig. 2). Negative \u0394\u03be displaces the contact pattern toward the inner edge of the gear tooth. Positive \u0394\u03be displaces the contact spot from the base of the tooth to the tip for the drive gear and in the opposite direction for the driven gear. 2. The length a\u03be of the major semiaxis in the instan taneous contact ellipse at tooth contact in the center of the engagement interval. 3. The angle \u03bb between the contact path (the trajec tory of the center of the instantaneous contact ellipse) and the perpendicular to the line of the driven gear at the calculation point (Fig. 2). 4. The maximum gap f on the interval of tooth engagement (the nonuniformity of torque transmis sion). 5. The modified roll coefficient Km. Synthesis yields the settings of the gear cutting machine required to produce the gear teeth. Recom mendations regarding the selection of all the relevant parameter values were made in [10]. Thus, the problem reduces to determining the set of parameters such that Eqs. (1)\u2013(5) are satisfied with maximum a\u03be. Synthesis and subsequent analysis of the gear engagement is based on the following assumptions",
            " At the center of engagement, the ellipse is defined by the lengths a\u03be and a\u03b7 of the major and minor semiaxes, while its area S is propor tional to their product (6) The maximum pressure \u03c3co (at the center of the instantaneous contact ellipse) is (7) The normal contact force Fn is related to the speci fied torque Mgiv at the gear shaft approximately as fol lows (8) where r is the mean pitch radius of the gear; \u03b1 is the profile angle; \u03b2 is the helical angle. The maximum contact pressure in good engage ment must act far from the edges of the teeth. As a rule, this is the case for tooth surfaces satisfying the follow ing kinematic requirements. The contact line AB passes through a specified point P, whose position is determined by the coordi S \u03c0a\u03bea\u03b7.= \u03c3co 3 2 Fn S .= Fn Mgiv r \u03b1 \u03b2coscos ,= nates \u0394\u03be and \u0394\u03b7 in the system M\u03be\u03b7 with its origin at the midpoint M of the lateral tooth surface of the driven gear (Fig. 2). By adjusting \u0394\u03be and \u0394\u03b7, the contact pat tern may be shifted over the tooth surface. With uniform gear rotation from its initial position in any direction, a gap is formed between the surfaces of the tooth pair. We assume that, in its initial position, there is contact of this tooth pair at calculation point P, and the ratio of the angular velocities is equal to the gear ratio of the tooth pair. The orientation of the ellipse is determined by the angle \u03bd (Fig. 2) between the tooth line and the major semiaxis of the ellipse. This angle, which does not vary greatly during engagement, may be estimated as (9) The following relationships may be used in devel oping an optimization algorithm. They were estab lished in [11, 12]. With increase in the major semiaxis a\u03be, the minor semiaxis a\u03b7 of the contact ellipse is reduced. Accord ing to Eq. (6), that produces fairly slow increase in area of the ellipse (proportional to ) and reduction in the contact pressure The minor semiaxis a\u03b7 of the contact ellipse declines with increase in the inclination \u03bb of the work ing trajectory (Fig. 2). This effect is only pronounced at \u03bb > 80\u00b0. The center of the contact ellipse is shifted along the working trajectory in the course of engagement. As a rule, the contact path is close a straight line. Its direc tion is determined by the inclination \u03bb with respect to the \u03bd axis (line AB in Fig. 2). The overlap coefficient \u03b5 and length of the contact path increase with increase in \u03bb. The length of the contact path (line AB in Fig. 2) and the overlap coefficient during engagement decline with increase in nonuniformity f of transmission. The minimum contact pressure in the engagement of conical gears with spiral angle \u03b2 \u2264 35\u00b0 corresponds to \u03b5 \u2264 2. With a localized contact pattern, when \u03b2 < 35\u00b0, the contact pressure is a maximum in single pair contact. When \u03b5 > 2, localized contact is only pos sible when \u03b2 > 35\u00b0. Other conditions being equal, increase in f will reduce a\u03b7 and increase the contact pressure. With increase in \u03b2, the inclination \u03bd of the ellipse\u2019s major semiaxis to the tooth line will increase, while the possible length of the contact ellipse declines",
            " In that case, we adopt the initial approximation \u03bb = 70\u00b0. The ratio c = a\u03be/b is assumed to be 0.33. (Here a\u03be is the length of the major semiaxis of the instantaneous contact ellipse, and b is the face width of the tooth face.) In the case where the width of the tooth face and the spiral angle are small, transverse motion of the instantaneous contact area is preferable. If possible, the required sum Y of the distances from the longitu dinal edge is ensured by increasing the nonuniformity f and thereby reducing the length of the contact path AB (Fig. 2). Only if that is impossible, because it vio lates Eq. (5), is y further increased by increasing \u03bb. When \u03b2 is small (\u03b2 < \u03c0/6), localized contact may be obtained when \u03bb is close to 30\u00b0. This value of \u03bb is adopted as the initial approximation. At large \u03b2, localization of the contact pattern with instantaneous contact areas of considerable length is only possible at large \u03bb. We may adopt \u03bb = 50\u00b0 as the initial approximation. The permissible relative length of the contact area declines with increase in \u03b2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000612_iecbes.2014.7047531-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000612_iecbes.2014.7047531-Figure3-1.png",
        "caption": "Figure 3. The interconnected bodies of a kicking leg from overhead view presented as two bars mechanism.",
        "texts": [
            " However, each segment length was calculated based on the ratio of segment length to the human height where the measurement of both legs were assumed to be equal [8]. Consequently, the measured values of the leg length showed that there were differences between the right and left legs as shown in Table II. The results from the calculations are presented in Table V. For the analysis of the forces acting on the kicking leg, the kinetics of rigid bodies analysis were applied by applying the general equation of motion for systems of interconnected bodies. Fig. 3 shows the interconnected bodies representing the kicking leg of the subject at an impact from an overhead view for the theoretical calculation. Fig. 3 shows the free-body diagramme of a kicking leg for the thigh and shank segment only. The feet segment was not analysed as the dimension of the foot was small compared to the shank and thigh segment. But, the impact force direction was based on the ankle\u2019s angle at impact, showing that the foot was dorsi-flexed at an angle during the impact. The angle was the flexion angle of the hip while was the knee\u2019s flexion angle. II. The velocity and acceleration diagrammes For the kinematical analysis, the methods of velocity and acceleration diagrammes were utilised also by representing the kicking leg as a two-link mechanism of thigh and shank segment only as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002380_transjsme.15-00563-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002380_transjsme.15-00563-Figure2-1.png",
        "caption": "Fig. 2 Overview image about Foldable-Multi-Copter",
        "texts": [],
        "surrounding_texts": [
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n2. \u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u958b\u767a\n2\u00b71 \u554f\u984c\u63d0\u8d77\n\u672c\u7814\u7a76\u3067\u306f\uff0cUAV\u3068\u3057\u3066\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u7528\u3044\u308b\uff0e\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u8907\u6570\u306e\u30d7\u30ed\u30da\u30e9\u306b\u3088\u308a\u63a8\u529b\u3092\u5f97\u308b\u3053\u3068\u3067\u6d6e \u4e0a\u3057\uff0c\u305d\u308c\u305e\u308c\u306e\u30d7\u30ed\u30da\u30e9\u306e\u63a8\u529b\u6bd4\u3092\u5236\u5fa1\u3059\u308b\u3053\u3068\u3067\uff0c\u6a5f\u4f53\u306e\u59ff\u52e2\u5236\u5fa1\u3092\u884c\u3046\uff0e\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u56fa\u5b9a\u7ffc\u6a5f\u3068\u6bd4 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\u63a2\u67fb\u7528\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u958b\u767a\u3067\u306f\uff0c\u3067\u304d\u308b\u9650\u308a\u5927\u304d\u306a\u30d7\u30ed\u30da\u30e9\u3092\u5229\u7528\u3059\u308b\u3053\u3068\u304c\u671b\u307e\u3057\u3044\uff0e\n\u4e00\u65b9\uff0c\u73fe\u5728\u5c4b\u5185\u3067\u5229\u7528\u53ef\u80fd\u306a\u96fb\u52d5\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u6700\u5927\u3067\u3082 30\u5206\u7a0b\u5ea6\u3057\u304b\u98db\u884c\u3067\u304d\u305a\uff0c\u63a2\u67fb\u5730\u70b9\u306b\u5230\u9054\u3059\u308b\u307e \u3067\u306b\uff0c\u96fb\u6c60\u5927\u304d\u304f\u6d88\u8017\u3059\u308b\u53ef\u80fd\u6027\u304c\u8003\u3048\u3089\u308c\u308b\uff0e\u3053\u306e\u554f\u984c\u306b\u5bfe\u5fdc\u3059\u308b\u305f\u3081\uff0c\u6211\u3005\u306e\u30b0\u30eb\u30fc\u30d7\u3067\u306f\uff0c\u5efa\u7269\u4fb5\u5165\u5730\u70b9 \u304b\u3089\u63a2\u67fb\u5730\u70b9\u307e\u3067\uff0cUAV\u3092 UGV\u306b\u642d\u8f09\u3057\u3066\u904b\u642c\u3057\uff0c\u63a2\u67fb\u5730\u70b9\u3067 UAV\u3092\u7528\u3044\u305f\u63a2\u67fb\u3092\u5b9f\u65bd\u3059\u308b\u30b7\u30b9\u30c6\u30e0\u3092\u63d0\u6848 \u3057\u305f\uff0e\u3055\u3089\u306b\uff0c\u5b9f\u8a3c\u8a66\u9a13\u3092\u901a\u3058\uff0c\u305d\u306e\u30b7\u30b9\u30c6\u30e0\u306e\u6709\u7528\u6027\u3092\u793a\u3057\u305f (Michael et al., 2012)\uff0e\u3057\u304b\u3057\u306a\u304c\u3089\uff0cUGV\u304c\u63a2 \u67fb\u5730\u70b9\u307e\u3067\u8d70\u884c\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u30c9\u30a2\u3084\u5eca\u4e0b\u306a\u3069\u306e\u72ed\u9698\u7a7a\u9593\u3092\u901a\u308a\u629c\u3051\u308b\u5fc5\u8981\u304c\u3042\u308a\uff0cUGV\u306b\u642d\u8f09\u3059\u308b\u30de\u30eb\u30c1\u30b3\u30d7 \u30bf\u306e\u30b5\u30a4\u30ba\u3092\u5927\u304d\u304f\u3059\u308b\u3053\u3068\u304c\u3067\u304d\u306a\u3044\uff0e\n\u305d\u3053\u3067\uff0c\u672c\u7814\u7a76\u3067\u306f\uff0cUGV\u306b\u642d\u8f09\u6642\u306b\u306f\u30b3\u30f3\u30d1\u30af\u30c8\u306b\u53ce\u7d0d\u53ef\u80fd\u3067\u3042\u308a\uff0c\u98db\u884c\u6642\u306b\u306f\u5927\u304d\u306a\u30d7\u30ed\u30da\u30e9\u30b5\u30a4\u30ba\u3092\u78ba\n\u4fdd\u3067\u304d\u308b\uff0c\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u63d0\u6848\u3092\u884c\u3046\u3053\u3068\u3068\u3057\u305f\uff0e\n2\u00b72 \u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u8a2d\u8a08\n\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u30b3\u30f3\u30d1\u30af\u30c8\u306b\u53ce\u7d0d\u3059\u308b\u624b\u6cd5\u306f\u3044\u304f\u3064\u304b\u5b58\u5728\u3059\u308b\u304c\uff0c\u672c\u7814\u7a76\u3067\u306f\uff0c\u30d7\u30ed\u30da\u30e9\u3084\u30e2\u30fc\u30bf\u3092\u56fa\u5b9a\u3059\u308b \u8155\u90e8\u306b\u30d2\u30f3\u30b8\u3092\u53d6\u308a\u4ed8\u3051\uff0c\u6298\u308a\u305f\u305f\u3080\u65b9\u5f0f\u3092\u63a1\u7528\u3059\u308b 2\uff0e\u3053\u306e\u3088\u3046\u306a\u6298\u308a\u305f\u305f\u307f\u65b9\u5f0f\u3092\u63a1\u7528\u3059\u308b\u3053\u3068\u3067\uff0c\u683c\u7d0d\u6642\u306b \u306f\u6a5f\u4f53\u306e\u6700\u5c0f\u5e45\u304c\u5c0f\u3055\u304f\u306a\u308a\uff0c\u672c\u7814\u7a76\u304c\u5bfe\u8c61\u3068\u3059\u308b\u5ba4\u5185\u63a2\u67fb\u306b\u304a\u3044\u3066\uff0cUGV\u306b\u642d\u8f09\u3057\u3066\u306e\u72ed\u9698\u90e8\u901a\u904e\u304c\u53ef\u80fd\u3068\u306a \u308b\uff0e\u8155\u90e8\u3092\u6298\u308a\u305f\u305f\u3080\u624b\u6cd5\u306b\u95a2\u3057\u3066\u306f\uff0c\u3059\u3067\u306b\u8907\u6570\u306e\u88fd\u54c1\u304c\u5e02\u8ca9\u3055\u308c\u3066\u304a\u308a\uff0c\u904b\u642c\u6642\u306e\u7701\u529b\u5316\u3068\u3057\u3066\u6709\u52b9\u3067\u3042\u308b \u3053\u3068\u304c\u78ba\u8a8d\u3067\u304d\u308b\u304c\uff0c\u5e02\u8ca9\u54c1\u306f\uff0c\u4e00\u822c\u306b\uff0c\u98db\u884c\u524d\u306b\u4eba\u529b\u3067\u6a5f\u4f53\u306e\u5c55\u958b\u4f5c\u696d\u3092\u884c\u3046\u5fc5\u8981\u304c\u3042\u308b\uff0e\u305d\u3053\u3067\uff0c\u672c\u7814\u7a76\u3067",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u306f\u3053\u308c\u306b\u52a0\u3048\uff0c\u8155\u90e8\u306e\u6298\u308a\u305f\u305f\u307f\u65b9\u5411\u3068\u30d2\u30f3\u30b8\u69cb\u9020\u306b\u7740\u76ee\u3059\u308b\u3053\u3068\u3067\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306b\u3088\u308b\u81ea\u5df1\u5c55\u958b\u304c\u53ef\u80fd\u306a\u69cb \u9020\u3092\u63d0\u6848\u3059\u308b\uff0e\u63d0\u6848\u3059\u308b\u69cb\u9020\u3067\u306f\uff0c\u5c55\u958b\u30fb\u56fa\u5b9a\u306e\u305f\u3081\u306e\u98db\u884c\u306b\u76f4\u63a5\u95a2\u4fc2\u7121\u3044\u30a2\u30af\u30c1\u30e5\u30a8\u30fc\u30bf\u304c\u4e0d\u8981\u3068\u306a\u308a\uff0c\u6a5f\u4f53\u91cd \u91cf\u3092\u6291\u3048\u3089\u308c\u308b\uff0e\n\u63d0\u6848\u306b\u57fa\u304d\u8a2d\u8a08\u3057\u305f\u30d2\u30f3\u30b8\u90e8\u306e\u8a73\u7d30\u3092\u56f3 3\u306b\u793a\u3059\uff0e\u3053\u306e\u30d2\u30f3\u30b8\u306f\u30b9\u30e0\u30fc\u30ba\u306b\u53d7\u52d5\u52d5\u4f5c\u3059\u308b\u305f\u3081\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u304c \u52d5\u4f5c\u3057\u3066\u3044\u306a\u3044\u5834\u5408\uff0c\u91cd\u529b\u306b\u3088\u308a\uff0c\u8155\u90e8\u306f\u4e0b\u306b\u6298\u308c\u66f2\u304c\u308b\uff08\u56f3 3-\u5de6\uff09\uff0e\u4e00\u65b9\uff0c\u30d7\u30ed\u30da\u30e9\u306e\u63a8\u529b\u65b9\u5411\u3068\uff0c\u8155\u90e8\u306e\u5c55\u958b \u65b9\u5411\u3092\u4e00\u81f4\u3055\u305b\u308b\u3053\u3068\u306b\u3088\u308a\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u3092\u56de\u8ee2\u3055\u305b\u3066\u63a8\u529b\u3092\u5f97\u308b\u3053\u3068\u3067\uff0c\u8155\u90e8\u304c\u81ea\u7136\u306b\u5c55\u958b\u3059\u308b\uff08\u56f3 3-\u4e2d\uff09\uff0e \u8155\u90e8\u306f\uff0c\u6a5f\u4f53\u30d9\u30fc\u30b9\u90e8\u3068\u540c\u4e00\u5e73\u9762\u4e0a\u3068\u306a\u308b\u5730\u70b9\u3067\uff0c\u30b9\u30c8\u30c3\u30d1\u306b\u3088\u308a\u5c55\u958b\u304c\u5236\u9650\u3055\u308c\uff0c\u4e00\u822c\u7684\u306a\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u5f62\u72b6 \u3068\u306a\u308b\uff08\u56f3 3-\u53f3\uff09\uff0e\u8155\u90e8\u306e\u7a81\u8d77\u3068\u30b9\u30c8\u30c3\u30d1\u304c\u5d4c\u307e\u308a\u5408\u3046\u3053\u3068\u3067\uff0c\u30d2\u30f3\u30b8\u90e8\u306e\u904a\u3073\u306b\u3088\u308b\u8155\u90e8\u306e\u306d\u3058\u308c\u30fb\u632f\u52d5\u3092\u9632\u3050\uff0e\n\u306a\u304a\uff0c\u672c\u8ad6\u6587\u3067\u306f\u6298\u308a\u305f\u305f\u307f\u6a5f\u69cb\u306e\u6709\u7528\u6027\u3092\u691c\u8a3c\u3059\u308b\u3053\u3068\u3092\u76ee\u7684\u3068\u3059\u308b\u305f\u3081\uff0c\u6a5f\u4f53\u306b\u306f\u5b9f\u969b\u306e\u63a2\u67fb\u306b\u5229\u7528\u3059\u308b\n\u30bb\u30f3\u30b5\u3084\uff0c\u5916\u58c1\u3068\u306e\u63a5\u89e6\u9632\u6b62\u7528\u306e\u30d7\u30ed\u30da\u30e9\u30ac\u30fc\u30c9\u7b49\u306f\u642d\u8f09\u3057\u306a\u3044\uff0e\n2\u00b72\u00b71 \u8155\u90e8\u306e\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u306e\u691c\u8a3c\n\u8155\u90e8\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u3092\u691c\u8a3c\u3059\u308b\uff0e\u56f3 4\u306b\u53ef\u52d5\u8155\u90e8\u306b\u50cd\u304f\u529b\u3092\u793a\u3059\uff0e\u3053\u306e\u56f3\u306b\u304a\u3044\u3066\uff0c\u30d2\u30f3\u30b8\u306e\u56de\u8ee2\u4e2d\u5fc3\u3067\u3042\u308b\n\u70b9 Q\u56de\u308a\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u306e\u5f0f\u306f\u6b21\u5f0f\u306e\u3068\u304a\u308a\u3068\u306a\u308b\uff0e\u305f\u3060\u3057\uff0cl f \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u30d7\u30ed\u30da\u30e9\u4e2d\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0c l f g \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u53ef\u52d5\u8155\u90e8\u91cd\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0cm\u3092\u53ef\u52d5\u8155\u90e8\u306e\u8cea\u91cf\uff0cFm \u3092\u30d7\u30ed\u30da\u30e9\u63a8\u529b\uff0c\u03b8 f \u3092\u53ef\u52d5\u8155\u90e8\u306e \u6298\u308a\u305f\u305f\u307f\u89d2\u5ea6\uff0cg\u3092\u91cd\u529b\u52a0\u901f\u5ea6\u3068\u3059\u308b\uff0e\u306a\u304a\uff0c\u7c21\u5358\u306e\u305f\u3081\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306f\u8155\u90e8\u4e0a\u3067\u529b\u3092\u767a\u63ee\u3059\u308b\u3082\u306e\u3068\u3059\u308b\uff0e\n0 =\u2212l f gmgcos\u03b8 f + l f Fm (1)\n\u3053\u306e\u5f0f\u3092 Fm \u306b\u3064\u3044\u3066\u307e\u3068\u3081\u308b\u3068\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm = l f g\nl f mgcos\u03b8 f (2)",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3057\u3066\u3044\u308b\u5834\u5408\uff0c\u03b8 f = 0\u3067\u3042\u308b\u305f\u3081\uff0c\u53ef\u52d5\u8155\u90e8\u3092\u5b8c\u5168\u5c55\u958b\u72b6\u614b\u306b\u4fdd\u3064\u305f\u3081\u306e\u6761\u4ef6\u306f\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm \u2265 l f g\nl f mg (3)\n\u4e00\u65b9\u3067\uff0c\u6a5f\u4f53\u304c\u30db\u30d0\u30ea\u30f3\u30b0\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u6a5f\u4f53\u306e\u5168\u8cea\u91cf\u3092\u652f\u3048\u308b\u3060\u3051\u306e\u63a8\u529b\u304c\u5fc5\u8981\u3068\u306a\u308b\uff0e\u30db\u30d0\u30ea\u30f3\u30b0\u6642\uff0c4\u3064 \u306e\u30d7\u30ed\u30da\u30e9\u304c\u3059\u3079\u3066\u540c\u3058\u63a8\u529b\u3092\u767a\u63ee\u3057\u3066\u3044\u308b\u3068\u4eee\u5b9a\u3059\u308b\u3068\uff0c\u6a5f\u4f53\u5168\u4f53\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u304b\u3089\u30d7\u30ed\u30da\u30e9\u3042\u305f\u308a\u306e\u63a8\u529b Fm \u3092\u6c42\u3081\u308b\u3068\u6b21\u5f0f\u3067\u8868\u73fe\u3067\u304d\u308b\uff0e\u305f\u3060\u3057\uff0cmb \u3092\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u8155\u90e8\u4ee5\u5916\u306e\u6a5f\u4f53\u8cea\u91cf (\u30d9\u30fc\u30b9\u90e8\u8cea\u91cf)\u3068\u3059\u308b\uff0e\n4Fm = (mb +4m)g\nFm = ( 1 4 mb +m ) g (4)\n\u3053\u306e\u5f0f\u3068\uff0c\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3059\u308b\u305f\u3081\u306e\u6761\u4ef6\u5f0f (3)\u304b\u3089\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306b\u53ef\u52d5\u8155\u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u306a\u3044\u305f\u3081\u306e\u6761 \u4ef6\u304c\u6b21\u5f0f\u3068\u306a\u308b\uff0e( 1 4 mb +m ) g \u2265 l f g l f mg\nmb \u2265 4m ( l f g\nl f \u22121\n) (5)\n\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3067\u306f\uff0c\u53ef\u52d5\u8155\u90e8\u306e\u5148\u7aef\u306b\u6700\u91cd\u91cf\u7269\u306e\u30e2\u30fc\u30bf\u304c\u3064\u3044\u3066\u304a\u308a\uff0c\u305d\u306e\u70b9\u3067\u63a8\u529b\u304c\u767a\u63ee\u3055\u308c\u308b\uff0e\u305d\u306e \u305f\u3081\uff0cl f g < l f \u3068\u306a\u308a\uff0c\u5f0f (5)\u306e\u53f3\u8fba\u306f\u5e38\u306b\u8ca0\u3068\u306a\u308b\u3053\u3068\u304b\u3089\uff0c\u5e38\u306b\u6210\u308a\u7acb\u3064\uff0e\u3053\u308c\u3088\u308a\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u98db\u884c\u6642\u306b\u8155 \u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u3053\u3068\u306f\u7121\u3044\u3053\u3068\u304c\u8a3c\u660e\u3055\u308c\u305f\uff0e\n2\u00b72\u00b72 \u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u6642\u306e\u52d5\u4f5c\n\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306f\u3059\u3079\u3066\u306e\u30d7\u30ed\u30da\u30e9\u304c\u307b\u307c\u304a\u306a\u3058\u56de\u8ee2\u6570\u30fb\u63a8\u529b\u3092\u767a\u751f\u3057\u3066\u3044\u308b\u304c\uff0c\u79fb\u52d5\u306e\u305f\u3081 \u306b\u306f\u5404\u30d7\u30ed\u30da\u30e9\u3067\u63a8\u529b\u5dee\u3092\u4f5c\u308b\u3053\u3068\u3067\u6a5f\u4f53\u3092\u50be\u3051\uff0c\u6c34\u5e73\u65b9\u5411\u306e\u63a8\u529b\u3092\u5f97\u308b\uff0e\u307e\u305f\uff0c\u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u306f\uff0c\u5404\u30d7\u30ed\u30da\u30e9\u306b \u3088\u308a\u767a\u751f\u3059\u308b\u53cd\u30c8\u30eb\u30af\u3092\u5229\u7528\u3057\u3066\u5b9f\u73fe\u3059\u308b\uff0e\u305d\u306e\u305f\u3081\uff0c\u79fb\u52d5\u6642\u3084\u30e8\u30fc\u56de\u8ee2\u6642\u306b\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u3088\u308a\u3082\u63a8\u529b\u304c\u5c0f \u3055\u304f\u306a\u308b\u30d7\u30ed\u30da\u30e9\u304c\u5b58\u5728\u3059\u308b\u3053\u3068\u306b\u306a\u308a\uff0c\u524d\u8ff0\u306e\u8155\u90e8\u3092\u4fdd\u3064\u6761\u4ef6\u3092\u6e80\u305f\u3059\u3053\u3068\u304c\u3067\u304d\u306a\u3044\u5834\u5408\u304c\u5b58\u5728\u3059\u308b\uff0e\u7279\u306b"
        ]
    },
    {
        "image_filename": "designv11_64_0001214_ijhvs.2014.068101-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001214_ijhvs.2014.068101-Figure2-1.png",
        "caption": "Figure 2 Elevation of the wheelset viewed from rotating frame x direction",
        "texts": [
            " Velocity of the contact \u2018a\u2019 on the wheelset in the inertial frame XYZ is given by: a GXYZ XYZ XYZ XYZ V V Ga Ga \u2022 = + \u2126\u00d7 + (21) where GV is the velocity vector of the centre of mass, \u2126 ({ } )T\u03b8 \u03c9 \u03c8 is the total angular velocity of the wheelset and Ga is the displacement vector from the centre of mass of the wheelset to the contact point \u2018a\u2019 in frame XYZ. Ga is the velocity of contact point \u2018a\u2019 as seen from moving frame xyz. If ,[ ]ZR\u03c6 and ,[ ]XR\u03b8 represent the basic rotation matrices for rotation of \u03c6 about Z-axis and rotation of \u03b8 about X-axis, respectively, then the rotation matrix for coordinate transformation from frame xyz to frame XYZ is: [ ] cos sin cos sin sin sin cos cos cos sin . 0 sin cos R \u03c6 \u03c6 \u03b8 \u03c6 \u03b8 \u03c6 \u03c6 \u03b8 \u03c6 \u03b8 \u03b8 \u03b8 \u2212   = \u2212     (22) From Figure 2, displacement Ga in xyz frame is given as: { }0 00 .T xyzGa Ga a a= \u2212 (23) Using coordinate transformation matrix [R], Ga in XYZ frame is given as: [ ]{ }0 00 .T XYZ Ga R Ga a a= \u2212 (24) Angular velocity of the moving frame xyz is given as: { } . T xyz \u03b8 \u03c9 \u03c8\u2126 = (25) Using coordinate transformation matrix [R], \u2126 in XYZ frame is given as: [ ]{ } . T XYZ R \u03b8 \u03c9 \u03c8\u2126 = (26) A bond graph capsule is created to realise the contributions of the first two terms of equation (21) toward the velocity components of contact \u2018a\u2019 on the wheelset"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002417_1.4033694-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002417_1.4033694-Figure2-1.png",
        "caption": "Fig. 2 Definition of the coordinate systems",
        "texts": [
            " As the proposed methodology starts from only a known main rotor profile, the effect on performance caused by perturbations on the profiles can easily be calculated. Therefore, this methodology fits in an optimization procedure to find the optimal rotor profiles for a given application (with corresponding flow rate and pressure ratio). The proposed integral methodology starts with a model for the geometry. Subsequently, the rotor geometry, the meshing conditions, and the control volume calculations are discussed. 2.1 Rotor Geometry. Coordinate systems are defined as in Fig. 2, with the gate (subscript 1) and main (subscript 2) rotors on the left- and right-hand sides, respectively. C is the rotor center distance. The gate rotor rotates in clockwise direction with angular velocity x1, and the main rotor in counterclockwise direction with angular velocity x2. A right-handed coordinate system is attached to each rotor with the z-axis along the rotor axis and the x-axis perpendicular to both rotor axes. The reference plane, z\u00bc 0, is taken in the end-plane of the rotors. Inlet and outlet ports are at z\u00bc L and z\u00bc 0, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001186_icumt.2014.7002105-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001186_icumt.2014.7002105-Figure8-1.png",
        "caption": "Fig. 8. Aerosonde UAV.",
        "texts": [
            " Additionally the control output, aileron, will also employ an interval type 2 triangular membership function. These membership functions are given in Fig. 7. The linguistic values that describe the input linguistic variables and also the output are: Too Negative (NN), Negative (N), Zero (Z), Positive (P), Too Positive (PP). In Table 1 rule base for the fuzzy interference system is given. In this paper the verifications are done with an accurate nonlinear dynamic model of Aerosonde UAV [23] which is seen in Fig. 8. The system specification of this UAV is given in Table II. Its dynamic model is included in the Aerosim Aeronatical Simulation Blockset [24], which provides a complete set of tools for 6-dof generic UAV models in the MATLAB environment. Aerosonde has great flexibility to use plane dynamics with user defined command and control algorithms. It also enables to launch UAV from desired coordinates. In the simulation of the benchmark system, a PI controller with KP = 8.7266 10\u22123 and KI = 8.7266 10\u22124 is chosen"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001996_sii.2015.7404997-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001996_sii.2015.7404997-Figure2-1.png",
        "caption": "Fig. 2. a) is model of 2-link manipulator (initial angle: q = [\u2212\u03c0/6,\u22122\u03c0/3]) and (b) is model of 3-link manipulator (initial angle: q = [\u2212\u03c0/6,\u2212\u03c0/3,\u2212\u03c0/3]T).",
        "texts": [
            "(1) by the proposed method described in chapter 2\u223cchapter 4, using the simulation of 2-link and 3-link manipulators. These manipulator models are with small linkages, but they are suitable for checking whether the proposed algorithm is correct. The simulation environment was used by \u201dBorland C++ Builder Professional Ver. 5.0\u201d for program creation, and the \u201dOpen GL Ver. 1.5.0\u201d for display. And we set the numerical integration time of Runge-Kutta to 1.0 \u00d7 10\u22122 [sec], the friction coefficient of ground to K = 0.2 and ft = 0.2fn. Figure 2 shows the simulation of restraint motion by the 2-link and 3-link manipulators. As the physical parameters, we set the mass and length of each link are mi = 1.0[kg], li = 0.5[m], the viscous friction coefficient of each joint is Di = 3.0[N\u00b7m\u00b7s/rad], and we set the initial position as shown in Fig.2(2-link:q = [\u2212\u03c0/6,\u22122\u03c0/3],3-link:q = [\u2212\u03c0/6,\u2212\u03c0/3,\u2212\u03c0/3]T), this simulation was performed by varying an input torque \u03c4 [N ]. As shown in Fig.2, it is the motion of y-z plane in both of the 2-link and 3-link manipulators. As an example of the arbitrary input for 2-link manipulator, when two kinds of input \u03c4 = [\u22123, 3]T, \u03c4 = [3 sin \u03c0t/3, 3 sin \u03c0t/3]T are given, and the change of time of z coordinate 0z3 of the tip of link 2 and the angle q1, q2 of each link are shown in Fig.3 and Fig.4 respectively. Similarly, when two kinds of input \u03c4 = [3, 3, 3]T, \u03c4 = [\u22123 cos 2\u03c0t,\u22123 sin 2\u03c0t, 3 cos 2\u03c0t]T for 3-link manipulator are given, the change of time of z coordinate 0z4 of 3rdlink\u2019s tip and the angle q1, q2, q3 of each link are shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002454_1350650116652566-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002454_1350650116652566-Figure2-1.png",
        "caption": "Figure 2. Oil film thickness calculation diagram under dynamic loading.",
        "texts": [
            " Computation model of sleeve bearing based on dynamic loading conditions Oil film thickness equation of spiral oil wedge sleeve bearing The structure of spiral oil wedge sleeve bearing is shown in Figure 1. The bearing has three tilted spiral oil wedges in the circumferential direction and both ends of every oil wedge have oil feed holes 2 and oil return holes 1. The special structure of sleeve bearing makes the oil film thickness to be calculated in the cylindrical surface and eccentric circular surface.17 Figure 2 shows the oil film thickness calculation diagram under dynamic loading. O is the bearing center, O2 is the axis center that is denoted by (x, y), Oh are the calculation points of counter-clockwise rotation along x-axis, and is attitude angle, cos \u00bc x=e, sin \u00bc y=e. Therefore, the oil film can be expressed as follows h \u00bc c x cos \u00fe y sin cylinder surface h \u00bc c x cos \u00fe y sin \u00feR1 cos \u00fe e1 cos\u00f0 \u00de R cos \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 e1 sin\u00f0 \u00de R1 h i2r eccentric circular surface 8>>>>>>>>< >>>>>>>>: \u00f01\u00de where h is the oil film thickness, e is the eccentricity, e1 is the eccentricity of eccentric circular surface, c is the bearing radial clearance, R1 is the radius of circular surface, is the dimensionless circumference coordinates, is the angular position of eccentric circular surface, R is the bearing radius, x is the circumferential direction, and y is the radial direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000318_tia.2010.2070784-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000318_tia.2010.2070784-Figure7-1.png",
        "caption": "Fig. 7. Configurations with the SL winding in healthy and faulty conditions. (a) SL healthy. (b) SL-1. (c) SL-2.",
        "texts": [
            " In this section, some considerations are carried out about the arrangement of the two sets of three-phase windings. Both healthy and faulty operating conditions are taken into consideration. The DL windings are shown in Fig. 6, in which only one of the two three-phase windings is shown. In configuration DL-1, each of the two three-phase windings is placed in a well-defined part of the stator [Fig. 6(a)]. On the contrary, the coils of the two windings are alternated along all the stator circumference in configurations DL-2 [Fig. 6(b)] and DL-3 [Fig. 6(c)]. Similarly, the SL windings are shown in Fig. 7. In the SL winding, the coils are wound around every other tooth, and each coil should be formed by a double number of turns (however, it is not possible in the prototype considered here). Configuration SL-1 has the two windings concentrated in a part of the stator [Fig. 7(b)]. Configuration SL-2 has the two windings with alternated coils [Fig. 7(c)]. 1) Healthy Operating Conditions: Depending on the winding arrangements, it is possible to take advantage of the presence of the two separate three-phase windings. The two windings can be supplied with currents \u201cin phase,\u201d as shown in Fig. 8(a). This means that the same current feeds phases A and A\u2032 and similarly for phases B and B\u2032 and for phases C and C \u2032. Alternatively, they can be supplied with currents out of phase of 30 electrical degrees, as shown in Fig. 8(b). However, not all the winding arrangements allow such a second supply strategy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003111_icemi.2015.7494242-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003111_icemi.2015.7494242-Figure2-1.png",
        "caption": "Fig. 2. Three-stage WT gearbox with one planetary and two parallel stages.",
        "texts": [
            " Wind turbines with more than 450 kW rated power have integrated gearboxes with a planet gear and two normal stages or two planet gears and one normal stage. The planet gear allows the load to be shared out between the planets and reduces the load at any one gear interface. As result gearboxes can be made smaller and lighter, at the cost of increased. Most of large gear-driven wind turbines use a three-stage design with a complex configuration as showed in Fig. 2T4l . The first stage is the planetary gear, which is more complicated than the parallel shaft. The low speed shaft stage in Fig. 2 shows planetary box is composed of planet gears, sun gear, ring gear and carrier. The ring gear is fixed to the gearbox housing and is not a moving component. The teeth of ring gear are on the inside of the ring and mesh with the teeth of planets at aIl time. The planet carrier holds three planets and ensures that the proper distances of ring-planet and planet-sun center are maintained. The sun is located in the center of the planets. A designer of wind turbines gearbox gives sorne general principles on the reliability of the three stage gearbox as follows l5l : \u2022 The high speed parallel stage is found to be the most unreliable module"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003051_0954407016629517-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003051_0954407016629517-Figure1-1.png",
        "caption": "Figure 1. Two-degree-of-freedom vehicle model (the bicycle model).26",
        "texts": [
            "24 limited the side-slip phenomenon to steady-state conditions but, in the case of friction between the tyre and the road, they discussed the Stribeck effect and differentiated between the friction coefficient in the longitudinal direction and the friction coefficient in the lateral direction. In the article by Lukowski et al.,25 the results of simulations using linear and non-linear side-slip characteristics of tyres are given. In this paper, a simple two-degree-of-freedom vehicle model was applied. It is very often called the bicycle model. A schematic diagram of the model presented by Ren\u0301ski and Sar26 is shown in Figure 1. If a constant longitudinal velocity is assumed, the vehicle motion can be described by two coordinates: the displacement along the y axis (with respect to the vehicle axis system) and the yaw angle c (the rotation about the z axis) or by their two derivatives, namely the lateral velocity _y and the yaw velocity _c respectively. Descriptions of the notation in Figure 1 are given in Appendix 1. For the model presented in Figure 1, the side-slip angles a1 and a2 are caused not only by the slip angles of tyres but also by the deformations of the suspensions and the steering system elements. The equation of forces along the y axis and the equation of moments are Fby +Y1 cos d1 +Y2 cos d2 +Fy = 0 \u00f01\u00de and Mb +Y1 cos d1 l1 Y2 cos d2 l2 +Mz = 0 \u00f02\u00de respectively, where Mb is the moment of inertia resistance given by Mb = Jz \u20acc. For small angles d1 and d2, it can be assumed that cos d1 = 1 and cos d2 = 1. When this simplification is taken into account, the motion of the vehicle can be described by the system of equations m _x _c + \u20acy + Y1 + Y2 +Fy = 0 \u00f03\u00de Jz\u20acc + Y1l1 Y2l2 + Mz = 0 \u00f04\u00de In the case of steady-state cornering, it can be assumed that the lateral forces Y1 and Y2 are functions of the axle slip angles a1 and a2",
            " The lateral forces Y1 and Y2 acting on the front axle and on the rear axle respectively are estimated on the basis of the measured lateral acceleration ay and the measured yaw acceleration \u20acc according to Y1 = mayl2 l1 + l2 + Jz\u20acc l1 + l2 \u00f011\u00de and Y2 = mayl1 l1 + l2 Jz\u20acc l1 + l2 \u00f012\u00de respectively. With reference to the tyre, the slip characteristic describes the dependence between the lateral force and the side-slip angle of the tyre. However, in the case of the single-track vehicle model presented in Figure 1, the relations between the force Y and the slip angle a can be regarded as the force\u2013slip characteristics of the axles. The vertical load on the tyre and the camber angle of the wheel influence the side-slip characteristics.28 The influence of the camber angle on the vehicle motion has also been presented by Virz\u0131\u0300 Mariotti and Ficarra.29 Furthermore, the manoeuvring dynamics play a significant role. In many publications, the nonsteady-state conditions of a tyre are described by the von Schlippe\u2013Dietrich formula,30 which includes a variable, namely the so-called tyre relaxation length"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000867_ukci.2014.6930154-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000867_ukci.2014.6930154-Figure1-1.png",
        "caption": "Fig. 1. Diagram of the underactuated hovercraft\u2019s kinematic model",
        "texts": [
            " In Section III using differential flatness theory dynamic feedback linearization of the hovercraft\u2019s model is performed and a state feedback controller is designed. In Section IV the Derivativefree nonlinear Kalman Filter is used as a disturbance observer so as to estimate the non-measurable state variables of the vessel as well as additive disturbance inputs that affect the hovercraft\u2019s model. In Section V simulation tests are carried out to evaluate the performance of the hovercraft\u2019s nonlinear control scheme. Finally, in Section VI concluding remarks are stated. The sixth order state-space equation of the underactuated hovercraft model (Fig. 1) is given by ?\u0307? = \ud835\udc62\ud835\udc50\ud835\udc5c\ud835\udc60(\ud835\udf13)\u2212 \ud835\udc63\ud835\udc60\ud835\udc56\ud835\udc5b(\ud835\udf13) ?\u0307? = \ud835\udc62\ud835\udc60\ud835\udc56\ud835\udc5b(\ud835\udf13) + \ud835\udc63\ud835\udc50\ud835\udc5c\ud835\udc60(\ud835\udf13) ?\u0307? = \ud835\udc5f ?\u0307? = \ud835\udc63\u22c5\ud835\udc5f + \ud835\udf0f\ud835\udc62 ?\u0307? = \u2212\ud835\udc62\u22c5\ud835\udc5f + \ud835\udefd\ud835\udc63 ?\u0307? = \ud835\udf0f\ud835\udc5f (1) where \ud835\udc65 and \ud835\udc66 are the cartesian coordinates of the vessel, \ud835\udf13 is the orientation angle, \ud835\udc62 is the surge velocity, \ud835\udc63 is the sway velocity and \ud835\udc5f is the yaw rate. The hovercraft\u2019s model is also written in the matrix form: \u239b \u239c\u239c\u239c\u239c\u239c\u239c\u239d ?\u0307? ?\u0307? ?\u0307? ?\u0307? ?\u0307? ?\u0307? \u239e \u239f\u239f\u239f\u239f\u239f\u239f\u23a0 = \u239b \u239c\u239c\u239c\u239c\u239c\u239c\u239d \ud835\udc62\ud835\udc50\ud835\udc5c\ud835\udc60(\ud835\udf13)\u2212 \ud835\udc63\ud835\udc60\ud835\udc56\ud835\udc5b(\ud835\udf13) \ud835\udc62\ud835\udc60\ud835\udc56\ud835\udc5b(\ud835\udf13) + \ud835\udc63\ud835\udc50\ud835\udc5c\ud835\udc60(\ud835\udf13) \ud835\udc5f \ud835\udc63\ud835\udc5f \u2212\ud835\udc62\ud835\udc5f + \ud835\udefd\ud835\udc63 0 \u239e \u239f\u239f\u239f\u239f\u239f\u239f\u23a0 + \u239b \u239c\u239c\u239c\u239c\u239c\u239c\u239d 0 0 0 0 0 0 1 0 0 0 0 1 \u239e \u239f\u239f\u239f\u239f\u239f\u239f\u23a0 ( \ud835\udf0f\ud835\udc62 \ud835\udf0f\ud835\udc5f ) (2) or equivalently, one has the description \u02d9\u0303\ud835\udc65 = \ud835\udc53("
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002705_amm.846.199-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002705_amm.846.199-Figure1-1.png",
        "caption": "Fig. 1 8-DOF LRV model [5]",
        "texts": [
            " No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (#73967904, Iowa State University, Ames, USA-15/02/17,09:27:28) governing each constituent element. The underlying objective is to discretise the system into a sufficient number of DOFs to quantify the actual continuous media system in the frequency range up to some \u03c9n, which is the highest natural frequency of interest for the system. For the purpose of this study the 8-DOF model featured in Fig. 1 is used across two case studies. Firstly, the rotor-dynamic model consists of three mass stations, in particular, one centrally mounted disc and two flexible isotropic half-shafts each supported by a dynamically linear bearing. The central concentrated mass (disk) has transverse and polar moments of inertia whilst able to translate in the x-y plane, thus having 4-DOF, namely displacements x and y and angular displacements \u03b8x and \u03b8y). Additionally, each bearing is allowed planar displacements in the x and y directions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.22-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.22-1.png",
        "caption": "Fig. 10.22 The skin effect: (a) current i(t) induces flux \u03a6(t), which in turn induces eddy currents in conductor; (b) the eddy currents tend to oppose the current i(t) in the center of the wire, and increase the current on the surface of the wire",
        "texts": [
            "60) Appendix B contains tables of the mean lengths per turn of standard ferrite core shapes, as well as the areas of standard American wire gauges. Eddy currents also cause power losses in winding conductors. This can lead to copper losses significantly in excess of the value predicted by Eqs. (10.58) and (10.59). The specific conductor eddy current mechanisms are called the skin effect and the proximity effect. These mechanisms are most pronounced in high-current conductors of multi-layer windings, particularly in highfrequency converters. Figure 10.22a illustrates a current i(t) flowing through a solitary conductor. This current induces magnetic flux \u03a6(t), whose flux lines follow circular paths around the current as shown. According to Lenz\u2019s law, the ac flux in the conductor induces eddy currents, which flow in a manner that tends to oppose the ac flux \u03a6(t). Figure 10.22b illustrates the paths of the eddy currents. It can be seen that the eddy currents tend to reduce the net current density in the center of the conductor, and increase the net current density near the surface of the conductor. The current distribution within the conductor can be found by solution of Maxwell\u2019s equations. For a sinusoidal current i(t) of frequency f , the result is that the current density is greatest at the surface of the conductor. The current density is an exponentially decaying function of distance into the conductor, with characteristic length \u03b4 known as the penetration depth or skin depth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure1.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure1.17-1.png",
        "caption": "Fig. 1.17 a Volt\u2013ampere characteristic and power-volt characteristic of solar cells. b Volt\u2013ampere characteristic of power-load element",
        "texts": [
            " If we try to use the value KPM , as the scale value, we cannot obtain the pure relative form of expression (1.32). 0.0 0.2 0.4 0.6 0.8 1.0 0.0 0.2 0.4 0.6 0.8 K PM K PM K P K G A=1.25 A=1.05 1.5 Nonlinear Characteristics 19 The solid arrows demonstrate the correspondence of the characteristic points. So, we have the three characteristic regimes. Then, the problem of scales, relative expressions, and correspondence of running points (dash arrow) arises. 1.5.2 Characteristic Regimes of Solar Cells Let us consider volt\u2013ampere and power-volt characteristics of two different solar cells in Fig. 1.17a. These characteristics have the maximum power points PM ; ~PM with the different currents IM ;~IM and voltages VM ; ~VM [8]. There are also such characteristic values as SC currents ISC;~ISC and OC voltages VOC; ~VOC . The solid arrows show the correspondence of the characteristic points. So, we have the three characteristic regimes. Similar to Sect. 1.5.1, the problem of scales, relative expressions, correspondence of running points (dash arrow), and approximation arises [15, 16]. 1.5.3 Quasi-resonant Voltage Converter It is known the load resonant converters, for example, with the zero-current switching [17]. These converters regulate their output by changing the switching period TS. Simulation of ORCAD model shows that its load curve is almost 20 1 Introduction rectangular with maximum load power point PM and similar to Fig. 1.17a. So, we have the same problems to determine the regime parameters. 1.5.4 Power-Source and Power-Load Elements The power-source and power-load elements are known from [4, 18]. Their load characteristics have the typical form as shown in Fig. 1.17b. Analysis of the power-load element shows the two-valued voltage of a limited capacity voltage source; that is, points M; ~M. But for all that, the volt\u2013ampere characteristic of this power-load element has one-valued representation. On the other hand, taking into account the losses of real power-source and power-load elements, we get the two-valued volt\u2013ampere characteristic of these elements. So, we have some characteristic regimes, and the same problems to determine the regime parameters. 1.6 Regulated Voltage Converters 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003055_mechatronika.2014.7018238-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003055_mechatronika.2014.7018238-Figure13-1.png",
        "caption": "Figure 13. Scheme of bearing coupling and acting forces.",
        "texts": [
            " The approach to bearing modeling presented in [12] is used here. The bearing couplings are supposed to be discrete and are concentrated into contact points between bearing rollingelements and the outer race fixed with the housing. The housing is supposed to be rigid and it means some modification of the approach used in [12]. The presented methodology enables to model both radial and radi-axial bearings. The bearing model respects real number of rolling-elements uniformly distributed between the inner and outer race (see Fig. 13) and includes nonlinear contact forces which depends on the shaft center deflection in the bearing cage and on the radial direction which the force acts. Let us suppose the rolling-element j of the bearing I touches the outer race at the contact point , . The force , transmitted at this point depends nonlinearly on the rolling- element deflection \u0394 , according to the Hertz\u2019s contact theory in following way [13] , , \u0394 , , , , \u0394 , . (11) The designation ax belongs to axial deflections of rollingelements and corresponding axial forces arising in the radiaxial rolling bearings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001123_1056789514560916-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001123_1056789514560916-Figure1-1.png",
        "caption": "Figure 1. Shaft with multiple cracks.",
        "texts": [
            " The input data to ANN provided by frequency responses spectra was obtained by finite element, theoretical, and experimental method and a good agreement was found with the existing data. With the help of linear fracture mechanics theory and taking the strain energy release rate and stress intensity factor, the mode shapes, natural frequencies, and stiffness of the steel cracked cantilever shaft have been calculated as follows. Estimation of local flexibility of damaged shaft under axial and bending load A multi-cracked cantilever shaft of diameter D is represented in Figure 1. Two transverse surface cracks are presented at two locations L1 and L2 with depth b1 and b2 from the fixed end (Figure 3). at Stockholm University Library on August 26, 2015ijd.sagepub.comDownloaded from Both the cracks result in a coupling effect yielding both longitudinal and transverse motion of the shaft. An axial force F1 and bending force F2 are applied at free end of the shaft. Due to the presence of cracks, a local flexibility will be interposed with the order of 2 2 matrix. The geometry of cracked surface is shown in Figure 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001513_cjme.2015.0302.022-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001513_cjme.2015.0302.022-Figure1-1.png",
        "caption": "Fig. 1. Tractive rolling contact between steel ball and control wheel",
        "texts": [
            " The geometric size of the contact area is determined and some novel methods and ideas are used to analyze the stress state in the rolling contact area in a step-by-step manner. This is done to obtain analytic solutions of the stresses and determine the characteristics of each stress component. Tractive rolling contact between a steel ball and a control wheel is crucial to automatic detection equipment detecting steel balls, and is a typical example of three-dimensional non-conforming rolling contact between a sphere and a cone. The contact is depicted in Fig. 1. Under the normal force ,P\u00a2 the steel ball and bilateral cones of the control wheel are brought together. Under the driving torque, the steel ball rotates about the axis AB, thereby causing the control wheel to rotate about the fixed axis CD. It is assumed that P is a normal force between the sphere and the one-side cone, with its direction passing through the center of the sphere (point O\u00a2 in Fig. 1) and the initial contact point (point o in Fig. 1). The contact surface is generated between the sphere and one side of the cone under the normal force P, and is considered to be an ellipse based on Hertz theory. The contact area is exaggerated in Fig. 1, and its actual dimensions in engineering practice are quite small, based on which the coordinate system is established. Following are descriptions of how to determine the geometric dimensions of the contact area and the analysis formulas of the stress. The general elliptical shape of the contact area was proposed by Hertz based on experimental observation of the interference fringes. He also introduced the simplification of each of the contacting bodies being an elastic half-space, and of the force acting on a flat surface in the small elliptical contact area",
            " In rolling contact between a sphere and a cone, one side of the contact area close to the rolling direction is under compression and the other side is under tension. The stick region therefore shifts in the rolling direction and the offset distances of all the point are equal. It is only at the boundary of the stick region (y=0) that the offsets coincide with the contact boundary owing to the special geometric area of the ellipse as discussed in Eqs. (1) and (2). The elliptical contact area is defined on the x-y plane, where o is the origin and initial contact point, and the x- and y-axis coincide with a and b as discussed earlier in connection with Fig. 1. The y-axis is in the direction of the cone generatrix, and the sphere rolls in the negative direction of the x-axis. The reversed z-axis is perpendicular to the x-y plane and points inwards of the cone in the rectangular coordinate system. As shown in Fig. 2, c and d denote the semi-axes of the elliptical stick region, which has its midpoint at o' and an offset distance of s. Y ZHAO Yanling, et al: Analysis and Numerical Simulation of Rolling Contact between Sphere and Cone \u00b7524\u00b7 With regard to the distribution of the tangential force, POPOV[31] used similar steps and assumptions as those proposed by CARTER[12] to calculate the threedimensional stress state in the rolling contact area of the sphere and plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001874_jbbbe.24.97-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001874_jbbbe.24.97-Figure2-1.png",
        "caption": "Figure 2: Second configuration of IPMC actuated mechanism for fin deflection",
        "texts": [],
        "surrounding_texts": [
            "Two proposed configurations of fin actuation mechanism are shown in Fig. 1 and 2. In first configuration, IPMC actuators are directly applied to the fin. Actuators act in cantilever form, fixed to base at one end and free to deflect at other end attached to the fin. Downward actuation of the actuator results in negative angle of attack of the fin and upward actuation results in positive. Whereas in second configuration, fin is actuated by the actuator through a connecting link (Link 2). Detailed description of the configuration with actuation principle is presented in [1]."
        ]
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.17-1.png",
        "caption": "Fig. 4.17 Revolute joint connecting bodies i and j",
        "texts": [
            " If the time is smaller than final time, go to step (2), otherwise stop the kinematic analysis. A close observation of the Eqs. (4.44) and (4.45) shows that both expressions represent systems of linear equations, with the same leading matrix and different right-hand side vectors. Moreover, since both expressions share the same leading matrix, Jacobian matrix of the constraints, evaluated with the latest calculated configuration of the system, then this matrix only needs to be factorized only once during each step (Nikravesh 1988). Figure 4.17 shows two bodies connected by a revolute joint in planar joint, which is a pin and bush type of joint that constrains the relative translation between the two bodies i and j, allowing only the relative rotation. The kinematic conditions for the revolute joint require that two different points at the center of the pin, each one belonging to a different body, share the same position in space the times. This means that the global position of the point Pi on body i needs to be coincident with the global position of the point Pj on body j"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.13-1.png",
        "caption": "Fig. 4.13 Global and local components of a point Pi on body i",
        "texts": [
            "25) can be expressed as ux uy \u00bc cos/i sin/i sin/i cos/i uni ugi \u00f04:26\u00de and in a compact form results in u \u00bc Aiu0i \u00f04:27\u00de where u is the vector expressed in terms of global coordinates, u0i is the vector expressed in the local coordinate system and Ai represents the planar transformation matrix for body i, which defines the orientation of body-fixed coordinate system \u03bei\u03b7i with respect to the global coordinate system xy, and given by Ai \u00bc cos/i sin/i sin/i cos/i \u00f04:28\u00de In the present work, the body-fixed or local components of a vector are denoted by an apostrophe, such as u0i. The analysis presented above is only valid for rigid bodies, that is, bodies in which the distances among their particles do not change during the motion of the body. Therefore, each point in a rigid body is located by its constant position vector expressed in the body-fixed coordinate system. For instance, a point Pi on body i, as shown in Fig. 4.13, can be described by the position vector sPi and by the global position of the body center of mass ri, resulting that rPi \u00bc ri \u00fe sPi \u00bc ri \u00feAis0Pi \u00f04:29\u00de where Ai represents the transformation matrix given by Eq. (4.28) and s0Pi refers to the local components of point Pi. The location of point Pi with respect to body-fixed coordinate system is s0Pi \u00bc nPi gPi T \u00f04:30\u00de In the expanded form, Eq. (4.29) is expressed as xPi yPi \u00bc xi yi \u00fe cos/i sin/i sin/i cos/i nPi gPi \u00f04:31\u00de or, alternatively xPi \u00bc xi \u00fe nPi cos/i gPi sin/i \u00f04:32\u00de yPi \u00bc yi \u00fe nPi sin/i \u00fe gPi cos/i \u00f04:33\u00de The local and global components of sPi are related as sPi \u00bc Ais0Pi \u00f04:34\u00de Equations (4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.20-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.20-1.png",
        "caption": "Fig. 3.20 Double wishbone wheel suspension with disconnected kinematic loops",
        "texts": [
            " The numerical complexity of the solution of the constraint equations is comparable with the numerical complexity of the sparse method. Example: Double wishbone wheel suspension After disconnecting the joints G1 and G2, a system with a tree structure and nine relative joint coordinates bi serves as the auxiliary variables to the motion description. Assembly: one obtains three constraint equations from each joint, e.g.: a1 \u00fe a2 b\u00f0 \u00de \u00bc a3 b\u00f0 \u00de \u00fe a4\u00f0b\u00de: \u00f03:13\u00de The final balance for the degrees of freedom results in 9 Auxiliary variables 9 2 3 Constraints -6 DoF 3 (one isolated DoF (Fig. 3.20)) Principle: Kinematic loops are treated as kinematic transmission elements, which in the following are characterized as the so-called kinematic transformers (Hiller et al. 1986, 1986\u20131988). Constraints: The constraint equations, which have to be stated, are composed of two different parts: \u2022 the non-linear, local transmission equations inside of each kinematic loop, \u2022 the linear equations that couple the kinematic loops. In this method, complete multibody loops are isolated and treated as local transmission elements"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.7-1.png",
        "caption": "Fig. 4.7 Relative coordinates: a linear displacement; b angular displacement",
        "texts": [
            " As far as the \u201crelative coordinates\u201d is concerned, it can be said that this type of coordinates was primarily used in the development of the first general computer programs for mechanisms analysis (Paul and Krajcinovic 1970; Sheth and Uicker 1971). The relative coordinates, also denominated as \u201cjoint coordinates\u201d, define the position and orientation of a body with respect to a preceding body in a multibody system. In general, this type of coordinates is directly associated with the relative degrees-of-freedom allowed by joints that connect bodies. Relative coordinates can be associated with linear or angular displacements, as shown by s and \u03d5 in Fig. 4.7. For the four-bar mechanism of Fig. 4.8, the set of relative coordinates that define its configuration can be stated as q \u00bc /1 /2 /3f gT \u00f04:15\u00de in which the three variables \u03d51, \u03d52 and \u03d53 represent the angle of each body with respect to the x-axis. Since the four-bar mechanism has only one degree-of-freedom, then the three relative coordinates \u03d51, \u03d52 and \u03d53 are not independent, and it is necessary to write a set of two constraint equations. In general, when working with relative coordinates, these equations can be obtained from the closed kinematic chain that defines the configuration of the system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003196_cistem.2014.7076970-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003196_cistem.2014.7076970-Figure3-1.png",
        "caption": "Fig. 3 .Structure du machine avec le cryostat",
        "texts": [
            " Afin d'am\u00e9liorer la distribution de l'induction radiale, on utilise aussi un mat\u00e9riau ferromagn\u00e9tique entre les deux bobines. Le supraconducteur est ins\u00e9r\u00e9 entre les pi\u00e8ces ferromagn\u00e9tiques comme repr\u00e9sent\u00e9 sur la Fig. 2. Le fer guidera le champ magn\u00e9tique vers le centre de l'inducteur et donc r\u00e9duira le champ de fuite \u00e0 proximit\u00e9 des bobines. Pour simplifier le syst\u00e8me de refroidissement de cette machine, nous avons choisi d\u2019avoir un induit tournant et un inducteur supraconducteur fixe et immerg\u00e9 dans un liquide cryog\u00e9nique comme il est montr\u00e9 dans la Fig. 3. La structure d'inducteur propos\u00e9e est en mesure de fournir une machine \u00e0 deux p\u00f4les et peut \u00eatre assimil\u00e9 aux machines synchrones \u00e0 griffes. Cette machine dispose d'un couple \u00e9lectromagn\u00e9tique \u00e9lev\u00e9e en raison de l'augmentation de la densit\u00e9 de flux dans l'entrefer. L'\u00e9tude a \u00e9t\u00e9 r\u00e9alis\u00e9e en ajoutant la culasse statorique pour montrer son influence sur le champ magn\u00e9tique g\u00e9n\u00e9r\u00e9 par cet inducteur. La simulation est r\u00e9alis\u00e9e dans deux cas, le premier, lorsque la structure d'inducteur ne contient pas le fer entre les deux bobines autour de la plaque supraconductrice, le second lorsque du fer est ajout\u00e9 en vue d'am\u00e9liorer la distribution de l'induction radiale d'entrefer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002781_chicc.2016.7554058-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002781_chicc.2016.7554058-Figure4-1.png",
        "caption": "Fig. 4: Constant avoiding angle \u03b10 between the instantaneous moving direction of the robot and one of the two boundary rays",
        "texts": [
            " The algorithm (12) steers the motion direction vector a(t) of the robot to a switching surface on which the robot\u2019s velocity v(t) becomes equal to the vector vi(t) + l (h) i (t). Now consider a coordinates system moving with the velocity vi(t). In these coordinates, the obstacle i becomes stationary, and the robot is moving with the velocity l (h) i (t). Furthermore, the motion with the velocity l (h) i (t) is an obstacle avoidance strategy which keeps a constant avoiding angle \u03b10 between the instantaneous moving direction of the robot and one of the two boundary rays of the vision cone of the covering sphere of the obstacle i from the robot; see Fig. 4. In other words, the control law consists of steering toward the closer of the two edges of the enlarged covering sphere vision cone defined by (8) in the plane defined by the robot\u2019s velocity vector and the direction towards the centre of the covering sphere of the obstacle. Our proposed strategy consists of switching between the obstacle avoidance law (12) and the straight motion to the final destination point at the maximum speed: u(t) = 0; V (t) = Vmax. (13) Let C > 0 be a given number. Furthermore, introduce the constants ai > 0 as ai := Ri cos\u03b10 \u2212Ri \u2200i = 1, 2, "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002425_s11771-014-2021-5-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002425_s11771-014-2021-5-Figure1-1.png",
        "caption": "Fig. 1 Enveloping graph of arbitrary tooth profiles",
        "texts": [
            " LUO et al [17] proposed a two-step method to manufacture this tooth profile, the first step is to generate a modified tooth gear by using a standard hobbing cutter in order to obtain a minimal machining allowance, and the second step is to generate the cosine tooth profile on a three-axis CNC milling machine. The main aim of this work is to propose a slotting method by using the involute cutter for the cosine gear based on the conjugate theories of digital surface and gear meshing. As shown in Fig. 1, the tool C2, whose equation is 2 2( , , ) 0,f x y   is assumed to be stationary in the coordinate systems 2(O2, x2, y2). O2 is its origin. The origin of a second coordinate system 1(O1, x1, y1) is offset from the origin O2. In the initial position, the axes y1 and y2 are coincident, y1 is parallel with y2. Suppose the tool C2 transmits movement to coordinate system  1(O1, x1, y1) and the moving parameter is \u03c6. The relation motion A(\u03c6) between the tool and the coordinate system  1(O1, x1, y1) can be expressed as 1 2( ) ( ) ( )   A A A (1) where A1(\u03c6) and A2(\u03c6) denote the rotation matrix and translation matrix, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001911_gt2015-44068-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001911_gt2015-44068-Figure13-1.png",
        "caption": "Figure 13: Typical brush seal design parameters",
        "texts": [],
        "surrounding_texts": [
            "Bristle diameter is one of the most important design parameters, which directly affects BTF and bristle stress levels at pressurized/unpressurized rotor interference conditions as well as steady state case. Typical industrial application for selecting the appropriate bristle diameter strongly depends on the experience, and larger diameters are usually preferred for applications that operate under high pressure loads. In order to improve the pressure load capacity of the brush seal, multi-stage and multi-layer configurations have been tried [18, 19, 20, 21, 22, 23]. However, due to nonlinear pressure drop between seal stages for multi-stage applications and introduced bristle locking phenomenon in multi-layer seal applications, the improvement in pressure load capacities are limited. Maximum pressure load capacity for a brush seal is reported as 27 bar (Dinc et al. [24]). In this section, the effect of bristle diameter on BTF and stress levels under transient and steady state conditions will be analyzed through correlated CAE models. Since the axial and tangential spacing parameters change with the bristle diameter, the number of bristle rows also changes (if other parameters such as bristle density, cant angle etc. are kept constant). Unpressurized rotor interference-BTF. The effect of bristle diameter on BTF at unpressurized-nonrotating rotor interference conditions (free-state rotor rub) has been examined by running the cases detailed in Tab. 4. Free-state BTF (BTF under unpressurized rotor rub) change with bristle diameter is given in Fig. 14 at different rotor interference levels. Second moment of area for the bristle crosssection is a 4th order function of the bristle diameter, therefore required tip force for bending the bristles is higher for larger bristle diameters. Increase in free-state BTF with bristle diameter is more pronounced at high interference levels. 7 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Steady state (\u0394P=0.3 MPa)\u2013VM Stress. The effect of bristle diameter on bristle stress levels for a loaded seal (pressure load without rotor interference) has been examined by running the cases detailed in Tab. 5. Stress levels at the most critical section, which is the FH point of backing plate side bristles (where downstream side bristles touching the backing plate corner, Figures 10 and 11), are examined during characterization study."
        ]
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure7-1.png",
        "caption": "Figure 7. (a) Deflections of the springs from Example 1, and (b) natural modes of vibrations for the system from Example 1 without dampers.",
        "texts": [
            "comDownloaded from In order to emphasize some of the benefits of the approach presented earlier, let us show the alternative way of finding the natural frequencies. The springs are undeformed (unstretched) in the stable static equilibrium position around which the particle performs small oscillations (Figure 5) and, consequently, one needs to determine the deflection in the direction that is collinear with the direction of the spring in the static equilibrium position (see Appendix 1 for detailed explanations of this fact). Based on Figure 7(a), these deflections are found to be 1\u00bc x=2\u00fey ffiffiffi 3 p =2 , 2\u00bc x=2 y ffiffiffi 3 p =2 , 3\u00bc x \u00f034a c\u00de and the potential energy is V \u00bc 1 2 k1 2 1 \u00fe 1 2 k2 2 2 \u00fe 1 2 k3 2 3 \u00bc 15k 8 x2 ffiffiffi 3 p k 4 xy\u00fe 9k 8 y2 \u00f035\u00de The kinetic energy of the particle is T \u00bc m _x2 \u00fe _y2 =2, so that Lagrange\u2019s equations of the second kind for small oscillations d dt @T @ _q \u00fe @V @q \u00bc 0, q 2 x, y \u00f036\u00de yield the following differential equation 4m \u20acx\u00fe 15kx ffiffiffi 3 p ky \u00bc 0, 4m \u20acy ffiffiffi 3 p kx\u00fe 9ky \u00bc 0 \u00f037a; b\u00de Assuming the solution for motion in the form x \u00bc A cos ",
            "2 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3\u00fe ffiffiffi 3 p 2 s ffiffiffiffi k m r \u00bc 1:966 ffiffiffiffi k m r \u00f040a; b\u00de Comparing the expressions for the frequencies from equation (40a,b) with the one given by equation (33a,b), it is seen that they coincide. Now, equation (38a,b) can be used to calculate y x !\u00bc!1 \u00bc B A !\u00bc!1 \u00bc 15k 4m!2 1ffiffiffi 3 p k \u00bc2\u00fe ffiffiffi 3 p \u00bc3:732, y x !\u00bc!2 \u00bc B A !\u00bc!2 \u00bc 15k 4m!2 2ffiffiffi 3 p k \u00bc 2\u00fe ffiffiffi 3 p \u00bc 0:268 \u00f041a;b\u00de The corresponding modes shapes are given in Figure 7(b), where the one corresponding to !1 is plotted as a dashed-dotted line and the one corresponding to !2 as a dotted line, and the value of A is taken arbitrarily to be equal to unity. It is important to note that the angles between these two directions and the horizontal are 75o and 15o, which agrees with the position of two-element equivalent system of springs shown in Figure 6. However, unlike this procedure related to natural modes of vibration, which involves lengthy calculations, the one related to the equivalent system of springs performed previously is straightforward and mathematically tractable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure14-1.png",
        "caption": "Fig. 14. Electric field distribution around the energized parts inside the opening of the tower",
        "texts": [
            " In this case finite element solution of COMSOL MultiPhysics was used to determine both the electric potential and the electric field distribution in the vicinity of the arrangement. The first part of the simulation was the determination of the electric potential of the different elements of the arrangement. Fig. 13 shows the values of the electric potential during the most critical conditions while the sinusoidal waveform of the voltage reaches its peak value. The nominal (RMS) value of the system voltage was 400 kV. From the electric potential values electric field distribution can also been determined. Fig. 14 and Fig. 15 shows the electric field distribution of the arrangement and in the vicinity of the conductor car. 978-1-5090-5165-6/16/$31.00 \u00a92016 IEEE From the results it can be determined that during practical working conditions peak value of electric field in the air in the surroundings of the arrangement is about 55 kV/m. Calculated with a 20 kV/cm value as the dielectric strength of the homogenous electric field and considering the inhomogeneity factor as 10, safety factor regarding to the electrical stresses is above 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002748_ijmic.2016.075271-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002748_ijmic.2016.075271-Figure2-1.png",
        "caption": "Figure 2 Illustration of region-based task of an AUV: pipeline inspection",
        "texts": [
            " Using a set of parameter vector 2 ,\u0398 \u2208R the restoring forces can be completely characterised as ( ) ( )\u03b7G \u03b7 Z \u03b7= \u0398 (6) where 6 2( ) .Z \u03b7 \u00d7\u2208R Detailed expression for Z(\u03b7) is given in Appendix B. 3 Region tracking 3.1 Desired region description and error dynamics AUVs perform different tasks like surveillance of a particular area or boundary line, specimen study, etc. These kinds of operational tasks involve a region-based regulation of the AUV rather than conventional set-point regulation problem. Figure 2 shows an AUV moving towards a pipeline for inspection. To define the region tracking problem, we need to consider an objective function. Here, the objective function describing the spherical attractive region as follows: ( ) ( ) ( ) ( )2 22 2 1 1 1 0d d df e x x y y z z r= \u2212 + \u2212 + \u2212 \u2212 \u2264 (7) where position error can be defined as 1 x d y d z d e x x e e y y e z z \u2212\u23a1 \u23a4 \u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5= = \u2212\u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u2212\u23a3 \u23a6 \u23a3 \u23a6 (8) If the region be considered as a circle in 2D or a sphere in 3D and the radius tends to zero, then the circular region or the spherical region become a point and it is the well-known set point regulation problem"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure4-1.png",
        "caption": "Fig. 4 Upright boundary conditions.",
        "texts": [
            " The frozen domain is made by the hub bearing houses, the connections of the upright to the suspension wishbones and steering rod, by the connections between the caliper and the upright and by the six cylinders and pad supports in the brake caliper. This domain is fixed and is not involved in the optimization process. Design and frozen domains have been discretized with linear tetrahedral elements, the average mesh size was 2 mm. Aluminum alloy has been considered as reference material for both components. The upright is fixed to the ground through three spherical hinges that represent the connections with the two wishbones and the steering rod. Each hinge is rigidly connected to the respective attachments on the upright as shown in Fig. 4. The caliper is rigidly connected to the upright through the two connection points. A tangential braking force is uniformly applied on the pads supports. This force is given by the friction generated when the pads are pressed against the disc surface. A uniform pressure load is applied in the six cylinders of the caliper to model the hydraulic pressure acting in the caliper. Loads acting on the upright come from the tyre/road contact forces. The lateral force is applied at the contact patch that is rigidly connected to the upright centre as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002901_stab.2016.7541212-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002901_stab.2016.7541212-Figure2-1.png",
        "caption": "Fig. 2. Dimensionless angular speed.",
        "texts": [
            " Due to the symmetry of the problem, there exists a trivial stationary motion (autorotation) for which the axis of symmetry of the body is vertical, the center of mass of the body moves along the axis of symmetry with a constant speed V, angular speed of autorotation about the axis of symmetry is a constant value \u03a9. For this regime, angles of attack of all blades are similar \ud835\udefc\ud835\udc56 = \ud835\udefc. Let\u2019s study influence of a blade shape upon stability of the trivial autorotation regime. Introduce a dimensionless angular speed \ud835\udf14 = \ud835\udc5f\u03a9/\ud835\udc49 of the body, where r is a distance between the axis of symmetry and a center of a blade. The instantaneous angle of attack and satisfy the following relations [4]: \ud835\udc58(\ud835\udefc) + \ud835\udc61\ud835\udc54(\ud835\udefc\u2212 \ud835\udefd) = 0, \ud835\udf14 = \ud835\udc58(\ud835\udefc) (1) where \ud835\udc58(\ud835\udefc) = \ud835\udc50\ud835\udc59(\ud835\udefc)/\ud835\udc50\ud835\udc51(\ud835\udefc) is an aerodynamic quality, \ud835\udefd is a pitch angle of a blade. The Fig. 2 represents curves \ud835\udf14=\ud835\udf1b(\ud835\udefd) for two different shapes of blades: a circle flat plate (grey curve) and a rectangular flat plate with aspect ratio 8 [5] (black curve). Notice that dimensionless angular speed \ud835\udf1b is ambiguous function of \ud835\udefd when \ud835\udefd\u223c\ud835\udf0b/2 . The domain of ambiguousness is wider for the 978-1-4673-9997-5/16/$31.00 c\u20dd2016 IEEE rectangular plate than for the circle plate. For the rectangular plate this domain is marked via dashed lines (Fig. 2). In the mentioned domain, there are several autorotation regimes, some of them are unstable. The corresponding condition of stability is given by the relation: \ud835\udc58\u2032(\ud835\udefc) > \u22121/ cos2(\ud835\udefc\u2212 \ud835\udefd) [4]. Let\u2019s describe intervals of \ud835\udefd, for which oscillations of the axis of symmetry converge at an autorotation regime for cases of different blade shapes taking into account displacement R of the center of mass of the body, assuming that mass of the body is fixed. Small oscillations of the axis of symmetry are described via complex differential equations of the third order with complex coefficients: "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000384_978-3-642-24145-1_11-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000384_978-3-642-24145-1_11-Figure3-1.png",
        "caption": "Fig. 3. Stress distributions of composite sandwich coupons loaded in both a) the machinedirection and b) the cross-direction",
        "texts": [
            " The bonding moduli of the corrugated sandwich coupon were calculated by using the following function: 33 4/ HblE \u22c5\u22c5\u22c5\u0394= Where \u0394 denotes the slope of the load-displacement curve, l is the span length, b is the specimen width and H the specimen thickness. The Bending Moduli was calculated to be 232 MPa for the coupon in the parallel-direction and was 333 MPa for the coupon in the cross-direction. In order to simulate the mechanical behaviour of corrugated sandwich structure, a FE model was created using Abaqus (Figure 3). The FE simulation will reduce the development costs and accelerate the development of the optimised structure in the early stage of design. At this stage, the adhesion between the core and the face sheets was ignored. As it can be seen from Figure 3 that the stress concentration occured on the corrugation core part and the region where the core and the face sheet meet. The model was able to simulate the load-displacement curves at the initial loading stage. The comparisons between the experimental and the numerical results are shown in Figure 4. It can be seen that at the initial loading stage, the FE models could simulate the experimental results reasonably well. A lightweight glass fibre composite structure has been fabricated for applications where high bending strength and stiffness are needed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003244_s1068366616040073-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003244_s1068366616040073-Figure1-1.png",
        "caption": "Fig. 1. Schematic of device for measuring stresses on surface of tool\u2013billet contact.",
        "texts": [
            " In solutions obtained using the slip-line technique, the tangential stress \u03c4XZ on the contact surface is determined by the slope of the \u03b1 slip line at the point of its outcrop to the contact surface and can be calculated using known equations of Mohr and Levi circles [6]. To obtain a solution using the slip-line technique with account for the above-mentioned equations, we introduce the following definition: \u03bc\u03c4 = cos2\u03c6, (3) where \u03c6 is the slope of the \u03b1 slip line to the X axis. The force of rolling PX and the thrust force PZ were determined using a split tool (Fig. 1). When the billet passed the clearance in the tool in the course of rolling, the measuring cells on which one of the components of the tool rested measured an increase in the components of the total force, i.e., the force of rolling \u0420X and the thrust force \u0420Z. The displacement of the billet was measured using a highly precise scanning displacement probe. The method of the experimental determination of contact stresses is described in [7]. Contact stresses during hot CR carried out at the temperature \u0422 = 1470 K were measured in steel 45 billets with diameters of 20 mm and lengths of 40 mm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003161_j.ifacol.2016.10.542-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003161_j.ifacol.2016.10.542-Figure1-1.png",
        "caption": "Fig. 1 Experiment system configuration",
        "texts": [
            " The subjects were healthy adult men with no history of conditions that could have a particular effect on standing or sitting movements. The mean age of the subjects was 32.25 \u00b1 7.46 years (mean value \u00b1 standard deviation), the mean height was 13th IFAC/IFIP/IFORS/IEA Symposium on Analysis, Design, and Evaluation of Human-Machine Systems Aug. 30 - Sept. 2, 2016. Kyoto, Japan Copyright \u00a9 2016 IFAC 267 262 Yusuke Kato et al. / IFAC-PapersOnLine 49-19 (2016) 261\u2013265 172.38 \u00b1 6.68 cm, and the mean body weight was 63.18 \u00b1 7.77 kg. The configuration of the experiment is shown in Fig. 1. To reduce the effects of movements of the upper limbs during the experiment, the subjects folded their arms in front of their trunk and repeated 10 standing and sitting movements at intervals of 7 s, which were timed with a metronome. On this occasion, we measured body movement and muscle activity. Note that each subject performed the movements twice; thus, a total of 20 standing and sitting movements were measured. There were no particular instructions regarding movement speed or foot position during the standing and sitting movements",
            " Next, there is gradual extension of the knee joint, the angle of the knee enlarges, and the buttocks are separated from the seat of the chair. After the buttocks leave from the chair, retroflexion of the trunk occurs, enlarging the AFI. The angle of the knee undergoes further extension, after the buttocks have left from the chair, and the person reaches a standing position. 2016 IFAC/IFIP/IFORS/IEA HMS Aug. 30 - Sept. 2, 2016. Kyoto, Japan Yusuke Kato et al. / IFAC-PapersOnLine 49-19 (2016) 261\u2013265 263 172.38 \u00b1 6.68 cm, and the mean body weight was 63.18 \u00b1 7.77 kg. The configuration of the experiment is shown in Fig. 1. To reduce the effects of movements of the upper limbs during the experiment, the subjects folded their arms in front of their trunk and repeated 10 standing and sitting movements at intervals of 7 s, which were timed with a metronome. On this occasion, we measured body movement and muscle activity. Note that each subject performed the movements twice; thus, a total of 20 standing and sitting movements were measured. There were no particular instructions regarding movement speed or foot position during the standing and sitting movements",
            " 4) was determined when the ground reaction force, which was measured by the force plate on the seat of the chair, was minimum. According to Fig. 4, anteflexion of the trunk occurs first during standing movements, thereby reducing the AFI. Next, there is gradual extension of the knee joint, the angle of the knee enlarges, and the buttocks are separated from the seat of the chair. After the buttocks leave from the chair, retroflexion of the trunk occurs, enlarging the AFI. The angle of the knee undergoes further extension, after the buttocks have left from the chair, and the person reaches a standing position. Fig. 1 Experiment system configuration Fig. 2 Placement of sensors Fig. 3 Photograph of experiment environment 2016 IFAC/IFIP/IFORS/IEA HMS Aug. 30 - Sept. 2, 2016. Kyoto, Japan According to Fig. 4, the minimal AFI is reached approximately when the buttocks leave the chair; therefore, we calculated the difference between the time when the buttocks left the chair and the time when the AFI was at its minimum for each subject (hereafter, the time difference). The results are shown in Table 1 and the mean time difference of all subjects is 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.8-1.png",
        "caption": "Fig. 10.8 Configuration of (a) turning and (b) boring",
        "texts": [
            " Furthermore, the MRR can be considered as the removed material volume throughout the process divided by Tcut. For the case L LA: MRR \u03c0D2 f 4 \u00bc kDLfeedv 4 \u00f010:3\u00de For turning, the material sample with a diameter D is mounted on a lathing machine which rotates with N revolutions per minute, whereas the cutting tool approaches the outer material layer gradually with a feed distance per revolution (Lfeed) or a feed rate f along the sample axial direction and the total cutting length along the axial direction L. as shown in Fig. 10.8a. The maximum cutting speed v is then equal to \u03c0ND. There is an offset distance LA of the cutting head position away from the sample surface when the process starts. Here, though the symbol definitions are different from the drilling process, the cutting speed v and the cutting time Tcut are still Eqs. 10.1 and 10.2, respectively. N 276 10 Design for Manufacturing In the turning process, it is assumed that the thickness of the material layer being chipped off is t, which is often much thinner than material diameter D. If we further let the material diameter after the process equal D0 (\u00bc D \u2013 2t), then MRR can be expressed by MRR \u00bc \u03c0 D2 D02 f 4 ktLfeedv \u00f010:4\u00de Boring is another machining process with the lathing machine as configured in Fig. 10.8b. A basic boring operation enlarges the inner diameter D0 of a pipeshaped structure to a larger diameter D by moving a cutting tool into the pipe cavity with a feed speed f. Under a rotational speed N with a unit of rpm, the maximum required cutting speed (v) is then v \u00bc \u03c0ND, the feed distance per sample rotation (Lfeed) is f/N, and the thickness of cut t is D0 \u2013 D, which should be much thinner than the inner wall diameter. The length of cut is assumed to be L. The cutting tool should have an initial offset distance of LA"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002778_j.proeng.2016.07.095-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002778_j.proeng.2016.07.095-Figure2-1.png",
        "caption": "Fig. 2.Construction of a tooth curve",
        "texts": [
            " Construction of tooth rims of a gear and a wheel The normal module mn is standardized. It\u2019s initial in case of geometrical and strength calculations of the helical gear. Therefore, we make a contour of the helical gear tooth in the Plane which is perpendicular to the tooth curve. The construction of the tooth rim we start with a helix g, which specifies a direction of the tooth curve. We carry out the following actions in order a middle point of the helix could coincide with the pitch point P (Fig. 2): by the command Plane we specify over a distance of 50 mm from the plane of the right view a plane, parallel to it and a face of the gear blank; in the created plane we make a sketch of the circle of the helix base with a diameter, which is equal to the reference diameter, and by the command Helix and Spiral we set the parameters: height 100; constant pitch 2896,2225; clockwise; initial angle6,215\u00b0. The initial angle which sets distribution of the initial point O of the We make a base circle dvb1of the equivalent wheel (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure7.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure7.6-1.png",
        "caption": "Fig. 7.6 (a) The basic design of an MRI scanner. Typically, it includes a 24-inch-wide chamber with a patient table transfer a patient into the chamber. Inside the MRI chamber, there are a magnet, a main-field coil, three pairs of gradient coils, and RF coils. The MRI scanner is connected to a computer system for signal processing. (b) A closer look of the gradient coils",
        "texts": [
            " If such a gradient magnetic field is only applied in an extremely short period of time Tny (within a quarter period of the MR signal oscillation for the boundary voxel), then the phase change \u03c6ny of the MR signals for different y-positions should become \u03c6ny \u00bc \u03c9n y\u00f0 \u00deTny \u00bc \u03b3n \u2202Bn \u2202y y\u00fe Bnz Tny \u00bc \u03c6no \u00fe \u03b3n \u2202Bn \u2202y Tnyy \u00bc \u03c6no \u00fe Gyy, \u00f07:22\u00de where \u03c6no \u00bc BnzTny. This implies that we can induce MR signals from the crosssectional layer with phase shifts linearly increasing with their y-positions. Therefore, the application of y-gradient magnetic field with a limited period is called \u201cphase encoding.\u201d Afterward, the MR signals will then be collected during the subsequent frequency-encoding stage. In the practical implementation, the MRI machine is typically configured as shown in Fig. 7.6a. The subject is lying on the patient table fitted inside the scanner chamber. It normally requires a primary superconducting magnet, a gradient coil, a main-field coil, and an RF coil. The primary magnet is used to generate a background magnetic field and force to align the spin direction of nuclei in hydrogen atoms. There is a horizontal tube that runs through the magnet called a bore. The magnet is extremely powerful, and its strength is measured in either \u201ftesla\u02ee or \u201fgauss\u02ee (1 tesla \u00bc 10,000 gauss). For comparison, most MRI magnets use a magnetic field of 0.5\u20132.0 tesla, whereas the Earth\u2019s magnetic field is only 0.5 gauss. There are three different pairs of gradient coils in the MRI machine which are located around the main magnet (Fig. 7.6b). Each pair of the gradient coils is responsible for generating a magnetic gradient along the Cartesian direction (x, y or z). They are weaker than the primary magnet as they are used to generate the magnetic gradient terms as described in Eq. 7.20. These gradient coils allow specific and different parts of the body to be scanned by altering and adjusting the magnetic gradients, coordinating simultaneously with the background magnetic field strength. A set of multiple RF coils is also equipped in the MRI scanner chamber"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000783_s13344-014-0065-9-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000783_s13344-014-0065-9-Figure3-1.png",
        "caption": "Fig. 3. Diving mechanism.",
        "texts": [
            " In fact, the guidance system determines the desired depth considering the projection of the distance between the AUV and the target in the horizontal plane to keep the AUV in a trajectory near the straight line connecting the two. With the straight line always close by, the trajectory shall be reasonably short and unsafe areas shall be avoided. It is assumed that the AUV wants to travel from (x, y, z) to (xp, yp, zp). The angle between the joining line and the horizontal plane ( 0 ) is shown in Fig. 3 and calculated as: 1 0 2 2 tan ( ) ( ) p p p z z x x y y        . (14) It is desired that the AUV moves toward the target keeping this angle. So the desired depth at time t is   2 2 d 0tan ( ) ( )p p pz t z x x t y y t             . (15) By calculating the desired depth and having the desired heading angle from the LOS strategy, the 3D guidance system is completed and the commands are prepared to be applied to the autopilot systems. According to Eqs. (4) and (5), the depth control for the assumed model of the AUV is independent of the heading control"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.11-1.png",
        "caption": "Fig. 6.11 Turning operation",
        "texts": [
            " For the different dimensional movements of the 152 6 Common Manufacturing Process worktable, the longitudinal traverse handwheel moves it horizontally (left and right), the cross traverse handwheel moves the work in and out, and the vertical movement crank moves the worktable and also the knee and saddle vertically (up and down) in unison. Turning is the machining process used to cut tubular stock into revolved shapes with precise diameters, which depends on the depth of cutting, by rotating the workpiece in a turning machine at high speed. The configuration of a turning process is shown in Fig. 6.11. The workpiece is then cut by the fixed cutting tool and creates waste material called chips. Products manufactured by turning are rotational, usually axisymmetric, with different features, including grooves, tapers, and holes. Turning 6.3 Secondary Processes for Metals and Alloys 153 is also applied as a secondary process to add or remove features on parts that were pre-generated by other processes such as casting, forging, or drawing. In the turning process, a variety of operations can be executed on the workpiece to produce various shapes as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure11-1.png",
        "caption": "Figure 11 Proposed duplex mechanism (see online version for colours)",
        "texts": [
            " To solve P1, we propose a new manipulator with a rail at the centre of rotation of the joints; to solve P2, we employ a pulling mechanism. By employing both pushing and pulling forces, deadlocks can be reduced. To solve the problems of the previous manipulator, we developed a new manipulator for the duplex mechanism, which is shown in Figures 10 and 11. The new manipulator is operated the same way as the previous manipulator (Section 3.3) except for the pulling mechanism. Each link has a semi-circular cross-section. Hence, by connecting two manipulators to each other, the shape becomes circular (Figure 11). As shown in Figure 11, the two manipulators share one rail that is installed at the centre. The rail is fixed to one manipulator, while the pulleys to guide the rail are fixed to the other manipulator. By this mechanism, the centre of rotation of each joint corresponds to the centre of the manipulator. Thus, problem P1 in Section 4 is solved. In addition, the size of the duplex system is reduced from that of the conventional one. The locking mechanism was also improved to realise the semi-circular shape in Figure 12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002562_978-3-319-32023-6_14-Figure14.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002562_978-3-319-32023-6_14-Figure14.1-1.png",
        "caption": "Fig. 14.1 Illustration of the mechanism of TN-90 LCD. Transmission axes of polarizer and analyzer are orthogonal to each other. Bright state is obtained in absence of applied voltage, i.e., normal white mode; however, if the transmission axes of polarizer and analyzer are parallel to each other, dark state is obtained in absence of applied voltage, which is called normal dark mode",
        "texts": [
            " We start our presentations from bistable CLC devices, including the generation of bistability and materials; followed by some specific application of CLC for solar energy manipulation in virtue of selective reflection, and subsequently, a new generation of LCDs driven by photovoltaic devices incorporated into the display architecture are introduced; finally, novel light-driven CLC photonic devices capable of exploiting solar energy for operation are demonstrated, focusing on the wide tunable Bragg reflector and the judiciously designed and fabricated micro-size integrated devices through micro-fluidic technique. Twisted nematic (TN) mode has been widely adopted in high definition displays at present, especially the well-known TN-90 mode, i.e. LC molecules rotate 90\u00b0 across the cell thickness on going from one substrate to the other [1, 2]. Device structure of TN-90 is shown in Fig. 14.1. Alignment directions on the two substrates are perpendicular with each other, which forces the direction of LC director near one substrate to be perpendicular to that close to the other, therefore forming the 90\u00b0 twisted arrangement of LC. It should be noted that to lift the double degeneracy of 90\u00b0 twisted arrangement, the nematic LC host is doped with a small amount (generally < 1.0 wt%) of chiral agent with weak twisting power. Such TN cell is inserted between a pair of crossed polarizers where the polarization direction is either parallel (i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001565_1350650115575026-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001565_1350650115575026-Figure1-1.png",
        "caption": "Figure 1. Rough cylinder on plane contact geometry.",
        "texts": [
            "56\u201359 Slide-to-roll ratio in microEHL has a significant influence on the wear rate, as the wear rate decreases with increase in slide-to-roll ratio when the slide-to-roll ratio is relatively small and a drastic rise in wear rate when slide-to-roll ratio is further increased. A thorough review of literature indicates that tremendous emphasis has been laid down to investigate the micro-EHL behavior in the past. However, the micro-EHL contacts lubricated with power law lubricant has not yet been thoroughly investigated. Hence, in the present work attempt has been made to investigate the micro-EHL line contact lubricated with power law fluids. A simplified contact formed between a rough cylinder and a plane as shown in Figure 1 has been studied. The theoretical model for the microEHL has been solved using finite element method technique. The complete investigation on the performance of micro-EHL contact under varying operating conditions as those prevailing during the operations of girth gears and bearings used in horizontal and vertical roller mills has been provided. This study will provide vital information for the design engineers that can aid in selecting and designing the components. The constitutive equation for the power law fluid model is expressed in tensor notation, where the stress tensor and strain tensor are related according to41,43 \u00bc m @u @z n 1@u @z \u00bc @u @z \u00f01\u00de where \u00bc m @u=@z n 1 is the equivalent viscosity of the fluid"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001996_sii.2015.7404997-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001996_sii.2015.7404997-Figure11-1.png",
        "caption": "Fig. 11. Forces acting on each link and accelerations of each link at 1\u00a9 t = 0.10 [sec] are shown at (a) as for y-direction and (b) as for z-direction. Also, at 2\u00a9 t = 4.85 [sec] they are shown at (c) as for y-direction and (d) as for z-direction. 3\u00a9 and the time of t = 3.95 as 4\u00a9. For 3\u00a9, 4\u00a9, Fig.16",
        "texts": [
            "9 shows the time response of the ycomponent of the force 0f1 = [0f1x, 0f1y, 0f1z]T, 0f2 = [0f2x, 0f2y, 0f2z]T acting on each joint and the frictional force ft acting on tip link. Figure 10 shows the time profile of the z-component of the force 0f1, 0f2 and constraint force fn, and Table 1 shows the physical parameters, and Table 2 shows the initial value and converged value for equilibrium configuration. We set 1\u00a9 as the time of initial motion at t = 0.10, and 2\u00a9 as the time at t = 4.85 that are shown in Fig.9, 10. Noticing the time of 1\u00a9, 2\u00a9, Fig.11 represents the force acting on y, z directions and the direction of acceleration at those time points. As the force acting on link i, fiy , fiz is the y-component, z-component of the force acting on link i from link i \u2212 1, \u2212f(i+1)y , \u2212f(i+1)z is the y-component, z-component of the force acting on link i from link i+1, fn is reacting force, ft is frictional force, mg is gravity. Figure 11(a) represents the motion of hand at time 1\u00a9 that accelerates toward y-axis minus direction, Fig.11(c) represents that the motion is decelerating in the y-axis plus direction. Although, there appears an acceleration to the falling direction(z-axis minus direction) in Fig.11(b) at time 1\u00a9, the direction of acceleration is z-axis plus direction at time 2\u00a9 as in Fig.11(d), it is a deceleration motion. Concerning force acting on link 1 and link 2, Fig.9 shows the force acting in y-direction, Fig.10 shows the force acting in z-direction. In the case of 1\u00a9 (t = 0.10), the force becomes 0f1 = [0,\u22120.129, 10.252]T, 0f2 = [0,\u22120.120, 0.464]T ft = 0.093, fn = 9.324. Then the translational acceleration of the center of mass of link i based on \u03a30 by 0s\u0308i = 0Ri is\u0308i, and A\u00a90s\u03081 = [0,\u22120.009,\u22120.012]T, 0s\u03082 = [0,\u22120.027,\u22120.012]T can be derived. Furthermore we used 0f\u0302i = [0f\u0302ix, 0f\u0302iy, 0f\u0302iz]T to show all the external forces acting on link i at the time of t = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.20-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.20-1.png",
        "caption": "Figure 9.20 Space-vector diagram corresponding to Usph= 202 V, 1 pu; EPMph= 186 V, 0.92 pu; Is = 115 A, 1 pu; cos\u03c6= 0.95; and \u03b4s= 24.5\u00b0. The d- and q-axis currents are 0.11 and 0.994 respectively. The rotor symbol indicates embedded magnets. Tem= 0.92.",
        "texts": [],
        "surrounding_texts": [
            "For a PMSM in particular, correct analysis of the flux-linkage reference and the inductances is necessary. In the following examples, consider the responses of a real PM machine to the application of different control methods. Table 9.1 gives the nameplate values and relevant parameters for the subject machine. Several different vector diagrams are illustrated in the upcoming examples for nominal operating points corresponding to rated load torque. Saturation of the inductances is not taken into account. In each example, the basic value for the flux linkage is always determined with respect to the stator voltage. This corresponds well with the presumption of a 3 % reduced voltage. Stator pu current for the given d-axis and q-axis current components is i 2 iq 2 0.112 0.9942 1.0is d The rated impedance as a base value is calculated as follows. UsNph 202 ZN 1.756 \u03a9 IsNph 115 The rated inductance is 0.4546\u03c8 sN 2.795 mH 2IN 2 115 LN The pu inductance values in d- and q-axis are as follows. Ld 1.12 mH Ld;pu 0.415 LN 2.795 mH Lq 1.16 mH Lq;pu 0.415 LN 2.795 mH The load angle is calculated trigonometrically. 0.92 0.11 0.401\u03c8PM Ldidcos \u03b4s 0.9 \u03b4s 25.4\u00b0 \u03c8 s 0.97 The voltage vector us is from \u03c8PM under angle \u03b4s+ 90\u00b0 = 25.4+ 90= 115.4\u00b0 and the current vector under angle 90\u00b0 + cos 1iq/is= 90\u00b0 + cos 1 0.994= 90\u00b0 + 6.3\u00b0 = 96.3\u00b0. The phase angle between the voltage and current and the corresponding power factor are \u03c6 115.4 96.3 19.1 cos \u03c6 0.94s s The machine is designed to operate about at id= 0 control. Therefore, the internal induced voltage must be slightly low (0.92 pu). Furthermore, the designer has included some voltage reserve for the machine drive. A converter connected to a 400 V network easily supplies a phase voltage fundamental of 220 V RMS, which corresponds to 1.09 V pu with the base voltage fixed to the motor rated voltage in this case. The base voltage could also be selected to be the converter maximum voltage when Ub= 230 V corresponding to 1 pu. In that case, the 202 V rated voltage should correspond to 0.88 pu. EXAMPLE 9.8: Produce a vector diagram based on the Table 9.1 nameplate values if id= 0and iq= is= 1at the rated frequency100 Hz,1 pu,EPMph= 186 V, Is= 115 A,1 pu,and Lq= 0.415 pu. The stator resistive voltage drop can be ignored. SOLUTION: Figure 9.21 is the space-vector diagram. The pu inductances of the motor are rather low, which is a good result. Supply voltage is higher than in the previous example. The stator flux-linkage reference in the DTC has been raised to the operating point shown in the figure. The power factor is smaller, and the voltage reserve is very small, which is typical of id= 0 control. EPMph 186 2 263.04 0.418 Vs and 0.92 pu\u03c8PM 2\u03c0f N 2\u03c0 100 628.31 Figure 9.21 Space-vector diagram according to the id= 0 control at a rated torque and a rated frequency. Usph= 210 V, 1.04 pu when the resistive voltage drop is taken into account. EPMph= 186 V, 0.92 pu; Is = 115 A, 1 pu; cos\u03c6= 0.91; \u03b4s= 24.4\u00b0; and Te= 0.92 pu. 2\u03c82 0.92 0.415 1 2 1.01 puLsis\u03c8 s PM 0.92\u03c8PMcos \u03b4s 0.91 cos 10.91 \u03b4s 24.4\u00b0 \u03c8 s 1.01 The power factor cos \u03c6 is calculated from the base of phase angle \u03c6, which is now identical to the load angle, because the current space vector is perpendicular to the flux linkage \u03c8PM and the voltage space vector is perpendicular to \u03c8s. Referring to the space-vector diagram, the power factor is cos \u03c6 0.91 The electromagnetic pu power and torque are as follows. Pe usiscos \u03c6 1 1 0.91 0.91 pu usiscos \u03c6 1.0 1 0.91 Te 0.91 pu \u03c9s 1 The same motor can be driven using a unity power factor. Figure 9.22 shows the resulting vector diagram. The current vector points in the negative d-axis direction resulting in smaller stator flux linkage, because of the armature reaction. Moreover, the voltage level is lower. The figure clearly indicates that to obtain an appropriate stator current at the rated torque, the flux linkage produced by the PMs should be higher. Here, the supply voltage is low, and as a result, the actual current demand exceeds the rated value by a considerable margin. There is plenty of voltage reserve. Because machine saliency is low in this case, torque is driven mainly by the q-axis current (TPM= iq\u00d7\u03c8PM= 0.99\u00d7 0.92= 0.91 pu), and the reluctance torque, expressed as Trel = (Lmd Lmq)idiq or Trel= (0.401 \u2013 0.415)\u00d7 ( 0.57)\u00d7 0.99 0.01, is insignificant. With increased PM flux linkage, the motor could be made to operate with a low voltage reserve and high power factor at its nominal operating point. However, operating the motor with a small load or at no load at the rated speed would prove problematic. In this case, EPM= 225 V. At small loads, demagnetization would be necessary to ensure sufficient voltage reserve. Figure 9.23 illustrates a machine at its nominal operating point that has been designed using a different method. In practice, this machine requires more magnetic material or magnets with higher coercive force. A design of this kind is particularly adapted for machines that do not operate in field weakening and do not demand dynamic performance. Designing a high-power high-efficiency blower using this method would be appropriate, for example. Figure 9.24 depicts the drive of the original motor, in which the stator flux linkage and the flux linkage of the PMs are regulated to be of equal magnitude in the DTC supply. In general, this is a good compromise, if good dynamic performance is required. The machine in this example has significant voltage reserve, the power factor is good, and the motor current is slightly higher than rated. Overall, these examples reveal how important stator flux linkage is to PMSM drive performance. The controller in a DTC application must pay special attention to the selection of the stator flux-linkage reference value. In steady state, a DTC converter may well find the MTPA operation by adjusting the stator flux-linkage reference to a value where the current is minimized at a given power level. and the frequency can be raised to 1.67 times the rated frequency (a 1 : 0.6 ratio). Since EPM at the rated frequency (100 Hz) is only 186 V, even a higher frequency can be reached by raising the voltage from the no-load voltage of 186 V to 230 V, in which case the frequency is 230/186\u00d7 167 Hz= 206 Hz. If the maximum converter current (1.28 pu) is used to demagnetize the motor at a full 230 V phase voltage, the maximum frequency of the motor in field weakening becomes 206 Hz\u00d7 INinv/INmotor= 206 Hz\u00d7 147 A/115 A= 264 Hz. At this frequency, the stator flux linkage would be \u03c8PM \u2013 idLd= 0.418 Vs \u2013 147 2 A\u00d7 1.12 mH= 0.185 Vs (41 % of the stator flux linkage at the nominal operating point). If for some reason, the demagnetizing current of the inverter should disappear at this operating point, the stator flux linkage would increase to 0.418 Vs, a value determined by the PMs. Correspondingly, the terminal phase RMS voltage of the motor would increase to 264 Hz\u00d7 2\u03c0\u00d7 0.418/ 2= 490 V. The line-to-line voltage should be 849 V, and the peak voltage of the intermediate DC link should be 1200 V, which would inevitably destroy the capacitors. A 230 V stator phase voltage can be produced by the inverter, so running at no load with no voltage reserve, the maximum practical field weakening for the subject machine is ffw= 230/186 100 Hz= 123 Hz. voltage is 230 V, and the stator flux linkage is 0.196 Vs (0.196 Vs\u00d7 2\u03c0264 3/ 2/s= 400 V, which corresponds to the converter maximum output line-to-line voltage. In field weakening, some and, in some cases finally, all of the stator current is used to reduce stator flux linkage. Consequently, in the constant flux range, more stator current is needed for field weakening than to produce torque. Figure 9.25 shows how different voltages behave for this motor at different supply frequency ranges. Figure 9.25 shows that field weakening in a PM machine is a complex matter. The major problem is the fixed flux linkage \u03c8PM, which can quickly lead to overvoltage risk to the inverter in field weakening. In this example case, frequency can be raised from the rated value of 100 Hz to 160 Hz without risking the inverter, because of the machine\u2019s rather low no-load voltage. If the no-load voltage is closer to the maximum voltage supplied by the inverter, failure risk moves closer to the nominal operating point. Figure 9.26 offers a space-vector diagram that represents operation at a safe upper limit of field weakening (160 Hz) when the inverter operates at its own rated current and the stator current corresponds to a stator voltage of 230 V. The stator flux linkage is \u03c8 s = 0.323 Vs, 0.71 pu. Figure 9.26 High-power operating point of the motor at the inverter rated current for 160 % rated speed. Usph= 230 V, 1.14 pu; EPMph= 298 V, 1.47 pu; Is= 147 A, 1.28 pu; cos\u03c6= 0.987; \u03b4s= 34\u00b0;P= 97 kW, 1.39 pu; f= 160 Hz, 1.6 pu;T= 0.915; andTN= 0.84 pu. Therefore, the motor runs close to its rated torque even at 60 % overspeed, producing much more than its rated power. The rated current of the inverter, which is higher than the rated current of the motor, and the slightly low PM flux linkage of the motor makes this possible. EXAMPLE 9.10: Prepare a space-vector diagram of a PMSM using the parameters and rated values given in Table 9.1, but with the frequency increased to 160 Hz, Usph= 230 V, EPMph= 298 V, converter maximum current Ismax= 147 A, and 1.28 pu as shown in Figure 9.25. Calculate the current components, load angle, power factor, power, and electromagnetic torque. SOLUTION: Figure 9.26 showed the relationships between voltage, current, and fluxlinkage vectors. The pu values can be calculated as follows. 230 us 1.138 202 298 ePM 1.47 202 230 2V 0.323 \u03c8 0.323Vs \u03c8 0.71s 2\u03c0 s;pu160 1 0.455s 298 2V 0.419 0.419Vs\u03c8PM \u03c8PM;pu 0.92 160 1 0.4552\u03c0 s The current components are solved based on the space-vector diagram. 2 2 2\u03c82 0.92 0.401 0.415 2 0.712Lqiq id iq\u03c8PM Ldids where i2 i2 i2 1.2782 i2q s d d Solved this equation returns the id and iq values. id 0.86 iq 0.945 The load angle is calculated as before. 0.92 0.401 0.86\u03c8PMpu Ldpuid cos \u03b4s 0.821 \u03b4s 34\u00b0 \u03c8 spu 0.71 or Lqiq 0.415 0.945 sin \u03b4s 0.56 \u03b4s 34\u00b0 \u03c8 s 0.71 The angle \u03b1 between is and d-axis is given by iq 0.945 sin \u03b1 0.738 \u03b1 47.7\u00b0 is 1.278 And, according the vector diagram \u03b1 \u03c6 \u03b4s 90\u00b0 \u03c6 90\u00b0 47.7\u00b0 34\u00b0 8.316 cos \u03c6 0.989 From this, it is possible to calculate the pu power and torque and the electromagnetic pu torque. Pe usiscos \u03c6 1.14 1.278 0.989 1.44 Te usiscos \u03c6 \u03c9s 1.44 1.6 0.9"
        ]
    },
    {
        "image_filename": "designv11_64_0001377_s00542-014-2085-z-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001377_s00542-014-2085-z-Figure4-1.png",
        "caption": "Fig. 4 Fabricated magnetic diaphragm after XeF2 etching. a PDMS cured at 343 K (curing time 30 min). b PDMS cured at 298 K (curing time 24 h)",
        "texts": [
            " a solenoid valve controls filling and evacuation of the chamber with XeF2 by detecting the signal from a pressure sensor connected to the chamber. The average time for the etching process denoted in Fig. 3 (evacuation time \u0394t1, filling time \u0394t2, holding time \u0394t3) are 110, 70, 60 s, respectively. The average etching rate is 1.2 \u03bcm/min. PDMS, TFPM and silicon have different thermal expansion coefficients, which are compared in Table 1. The thermal effect in the process has a big influence on the fabrication result. Figure 4 shows the fabricated magnetic diaphragms at different PDMS curing temperatures. The standard curing temperature of 343 K and curing time of 30 min for PDMS caused small cracks on the upper right hand side of the TFPM. By modifying the curing temperature to 298 K and extending the curing time to 24 h, the damage on the TFPM can be avoided as a result of the reduced internal stress in the PDMS. however, in both samples shown in Fig. 4, part of the TFPM has become detached from the PDMS diaphragm. This delamination is caused by the heat of the reaction in the XeF2 etching process which is evaluated in the following section. 3.1 evaluation of the TFPM magnetic performance In order to evaluate the magnetic properties of the TFPM before and after XeF2 etching, the J\u2013h curves of the TFPM were measured using a vibrating sample magnetometer. The second quadrant of the J\u2013h curves is shown in Fig. 5. after sputtering, the TFPM had a magnetic remanence of 1",
            " When the maximum process pressure is below 80 Pa, the maximum temperature drops below 430 K, at which temperature the magnetic performance of the TFPM is unaffected. 3.3 evaluation of the mechanical properties of the PDMS diaphragm In order to evaluate the changes in the mechanical properties of the PDMS diaphragm after XeF2 gas etching, the current\u2013displacement curve of the diaphragm was experimentally measured after re-magnetization. The measurement setup is shown in Fig. 9. The magnetic diaphragm fabricated utilizing the 150 Pa XeF2 etching process, shown in Fig. 4b, was sandwiched between two acrylic plates with 7 mm diameter holes at their centers. Uncured PDMS was used to bond the diaphragm to the acrylic plates. The force on the TFPM was generated using an electromagnet (Zhi et al. 2013). a laser displacement sensor (lKG155, Keyence corp.) was used to measure the diaphragm displacement. To convert the current displacement curve of Fig. 11a to a force displacement curve, the electromagnetic force was simulated by electromagnetic field simulation software (Maxwell 3D, ansys Inc",
            " In the simulation model, the edge of the PDMS film is clamped and a uniform pressure is distributed in the central circular area, as shown in Fig. 10. The simulated result is compared with the experimental one in Fig. 11b, which shows there is good agreement between the simulated and experimental data, thus demonstrating that the Young\u2019s modulus of 2 MPa and Poisson\u2019s ratio of 0.45 for the PDMS used in the simulation are the same as the conventional values (Suzuki et al. 2011; lu and Zheng 2004). Thus, we can conclude that the PDMS retains its mechanical properties after XeF2 gas etching. For the magnetic diaphragm fabricated in Fig. 4b, according to the results of the evaluation, the TFPM retained its magnetic performance (recovered after re-magnetization); meanwhile, the PDMS layer retained its mechanical performance. however, delamination of the TFPM from the PDMS layer gives rise to a shift in the effective center of the actuator from the actual center, which would cause diaphragm applied to micro devices to be misaligned; thus this must be avoided during the fabrication process. as the delamination is generated by thermal stress due to the high temperature (450 K) during the XeF2 etching process, the XeF2 maximum etching pressure was decreased in order to reduce the process temperature"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003522_cefc.2016.7815930-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003522_cefc.2016.7815930-Figure1-1.png",
        "caption": "Fig. 1 Configuration of the 12/10 FSPM machine and water cooling system.",
        "texts": [
            "78-1-5090-1032-5/16/$31.00 \u00a92016 IEEE Index Terms\u2014Axial segments, coupled-fields, flux-switching, magnetic-thermal, permanent magnet, water cooling. I. INTRODUCTION In a flux-switching permanent magnet (FSPM) motor [1], a water cooling system can be located on the shell, as shown in Fig. 1. Thus the PMs directly contact with the heat source and the dissipater simultaneously, i.e. the armature coils and motor shell. It can be expected that the temperature distributions in one magnet varies greatly, both in radial and axial directions. Fig. 2 gives the temperature distributions obtained by 3D finite-element analysis (FEA) and experimental measurements. Usually, to count in the 3D temperature distributions, a synchronous coupled magnetic-thermal fields 3D FEA model is required, as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003111_icemi.2015.7494242-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003111_icemi.2015.7494242-Figure4-1.png",
        "caption": "Fig. 4. Structure ofthe gearbox.",
        "texts": [
            " When the test rig is instrumented with accelerometer, torque transducer, eddy current displacement transducers, tachometer, voltage and current transducers, health condition of gearbox and motor can be described. Using the mechanical test rig, condition monitoring system and fault diagnosis techniques can be developed. Both motor electrical and WT gearbox mechanical faults can be simulated on the test rig with and without load. The gearbox has two parallel-shaft helical gear stages and a planet gear as showed in Fig. 4. The input shaft is high speed shaft connected with AC asynchronous motof. The input gear has 18 teeth while big gear of first stage has 72 teeth. Small gear of tirst stage has 20 teeth and big gear of second stage has 90 teeth. The form of planet gear used on the test rig is same as WT gearbox: An interior toothed gear wheel with 67 teeth, three smaller toothed planet wheels with 25 teeth carried on a planet carrier, a centrally placed toothed sun gear wheel with 17 teeth. High \u2022\u2022\u2022\u2022\u2022\u2022 speed shaft Highspeed shaft st age (c) Lowspeed shaftst age The CMS monitoring function is based on robust sensor equipment for continuo us measurements and performs online evaluation by use of modern digital signal processing methods"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002683_978-3-319-11930-4-Figure2.15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002683_978-3-319-11930-4-Figure2.15-1.png",
        "caption": "Fig. 2.15 Design optimization of chamber section of test rig C. The design couples the weldedbellows aspect with conflat flanges to enable vibration isolation at high temperature under vacuum (Reproduced with permission from Rev. Sci. Instrum. 82, 105113 Copyright 2011, AIP Publishing LLC)",
        "texts": [
            " These results suggest that a weldedbellows assembly should be considered in the design optimization of the next test rig. Test rig C was built to accommodate larger diameter rotating assemblies and to enable operation at higher temperatures. This design incorporates conflat flange joints with the isolation aspects of the welded-bellows assembly used in test rig A. The new chamber, test rig C, is an extension of the chamber design of test rig B with inclusion of the welded bellows from test rig A. The singular assembly that was optimized for test rig C was the main chamber shown in Fig. 2.15. The chamber design has four aspects: two welded bellows, tubular interface attachments, tubing wall thickness, and chamber wall thickness. As presented in Fig. 2.15, a conflat flange joint is used at one end of the chamber to enable removal and installation of the rotating assembly. During a test, the rotating assembly may be heated to 600 C while the vibration response of the vibration table is measured over time. With the rotating assembly at 600 C, the exterior of the chamber may reach 200 C due to radiative heat transfer inside the chamber. Hard copper gaskets are required for all conflat joints due to 24 2 Vacuum Chamber Design the heating conditions of this test rig",
            " The rotating assembly is driven by a hermetically sealed drive assembly similar to the one used on test rig A in Fig. 2.4. Two flexible bellows are used in the new design to enable more test configuration flexibility. The new chamber is made from 304 stainless steel (304SS), and all permanent seams are welded with no long-term gasket joints. Fabrication with stainless steel is expensive, and therefore, the amount of 304SS used should be minimized. The cost of cooling the chamber will be minimized through optimization of geometry and heat conduction through the half nipple attachments in Fig. 2.15. The interface attachment tube walls conduct heat from the chamber to the interface attachments. The design requirement concerning the half nipple attachments is that they cannot exceed room temperature, approximately 20 C, to protect and isolate temperaturesensitive measurement equipment from heat damage and leakage. The conductive heat transfer through the attachment tubes will be minimized within the constraints that the tube section of the attachments be able to support a 25 N cantilever load with an end deflection less than 0.1 mm. This requirement applies equally to all four tubes so that any instrument or pump can attach at any interface flange. The front surface of the chamber in Fig. 2.15 must not deflect more than 3 mm, and the chamber walls around the circumference must not deflect more than 1 mm. It is desirable that the entire chamber system has a first natural frequency greater than 600 Hz or three times the anticipated rotational speed of the rotating assembly. The cost of materials, fabrication, and operation to be minimized is given as 2.6 Case Study Continued: Optimum Chamber Design 25 f X\u00f0 \u00de \u00bc Costmetal \u00fe Costwelding \u00fe Costoperation: \u00f02:1\u00de The cost of the stainless steel for the chamber and tube interfaces is assumed constant",
            " Constraints 1 and 2 are classified as size constraints and 3 through 10 may be classified as material and behavioral constraints. Expressions and references for material and behavioral constraints 3 through 10 are established in the fields of mechanical engineering and applied physics and are presented below. The mathematical form of the optimization problem is stated as, min f X\u00f0 \u00de \u00bc \u03c1 A1L1 \u00fe A3L3 \u00fe A2t2\u00f0 \u00deCostmaterial \u00fe \u03c0 2d1 \u00fe 8d3\u00f0 \u00deCostweld \u00fe h Costcooling \u00f02:2\u00de subject to the constraint expressions in Eqs. 2.3, 2.4, 2.5, 2.6, 2.7, 2.8, 2.9, 2.10, 2.11, 2.12, 2.13, 2.14, 2.15, and 2.16 for the chamber shown in Fig. 2.15. Constraints g1 through g3 relate to radial and hoop stress and deflection of the chamber: 26 2 Vacuum Chamber Design g1 \u00bc p d1 t1 \u03c3y 0 \u00f02:3\u00de g2 \u00bc p d1 2 2 Et1 1 v 2 cdif 0 \u00f02:4\u00de g3 \u00bc p d1 2t1 \u03c3y 0: \u00f02:5\u00de Constraints g4 and g9 account for the first bending mode natural frequency of the chamber section and for each attachment tube, respectively. Since the rotating device inside the chamber may have an unbalance, or could develop an unbalance during testing, it is good practice to design for at least 2 the frequency response of the chamber system to internal excitation: g4 \u00bc freq 6:93 2\u03c0 ffiffiffiffiffiffiffiffiffiffi EI1g wL1 3 s 0 \u00f02:6\u00de g9 \u00bc freq 1:73 2\u03c0 ffiffiffiffiffiffiffiffiffiffi EI3g wL3 3 s 0: \u00f02:7\u00de Constraints g5 and g6 relate to front plate stress and bending deflection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure5.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure5.4-1.png",
        "caption": "Fig. 5.4 Schematic of debonding whisker",
        "texts": [
            " Such energy change associated with the bridging process is a function of the bridging stress divided by traction, \u03c3b, and the crack opening displacement, dc, and is defined as Jcb \u00bc Z dmax 0 \u03c3bddc, \u00f05:10\u00de where dmax is the maximum displacement at the end of the zone. One can equate the maximum crack opening displacement at the end of the bridging zone, dmax, to the tensile displacement in the bridging brittle ligament at the point of failure: 126 5 Ceramics dmax \u00bc \u03b5 lf ldb, \u00f05:11\u00de where \u03b5f l represents the strain to failure of the whisker and ldb is the length of the debonded matrix-whisker interface (Fig. 5.4). The strain to failure of the whisker can be defined as \u03b5 lf \u00bc \u03c3 l f=E l, \u00f05:12\u00de where El is the elastic modulus of the reinforcing phase and \u03c3f l is the fracture strength of the whisker. On the other hand, the interfacial debonded length depends on the fracture criteria for the reinforcing phase versus that of the interface and can be defined in terms of fracture stress or fracture energy: ldb \u00bc kdbr l\u03b3l=\u03b3i, \u00f05:13\u00de where \u03b3l represents the fracture energy (unit, J/m2) of the bridging ligament; \u03b3i represents the fracture energy of the reinforcement-matrix interface; and kdb is a constant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001877_amm.808.193-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001877_amm.808.193-Figure5-1.png",
        "caption": "Figure 5 - (left) Defined 3D-geometry of the lattice structure; (right) Probe with reduced geometry",
        "texts": [
            " Regarding the above mentioned marks that appeared on the surface or the lattice structures of the probes, there is no noticeable effect on the mechanical properties. it is important to reduce the experimental set-up on less to test parameters and to begin with a detailed and closer look on the material science. Microstructures that result from the material properties and the SLM process might provide information about fraction properties. Followed by a closer look on the basics of material science it is important to reduce the firstly defined 3Dgeometry of the lattice structure (fig.5 left) to a simplified 2D structure. This choice is necessary, because it is not sure if the complex lattice structure yet has an effect on the mechanical properties. To eliminate any disruptive factors and to follow the achieved results, the cross sectional area (with the major weakness) of the lattice structure is kept and a new probe is defined out of that shape (fig.5 right). The same applies to the probe with the coating. These reductions should answer the main issue about the achieved discrepancies and allow an easier micrograph. Whereas the main focus was on the geometry and coating of the lattice structure it has changed to the strained cross sectional area and material science. In reference to the first series of tests, results and interpretations following questions need to be answered to define the work schedule. The most important and eye-catching findings of the first series of tests are the discrepancies of the calculated tensile strengths"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000395_978-94-007-4132-4_8-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000395_978-94-007-4132-4_8-Figure7-1.png",
        "caption": "Fig. 7. Technical drawing of Hoecken\u00b4s six-link dwell mechanism showing details for its manufacturing",
        "texts": [
            " 6), we find out at once that Hoecken\u00b4s choice corresponds with the number 1 of Hain\u00b4s catalogue: there is a four-link crank-rocker with a two-bar E-F coupled to it. It is very interesting to discover that Hoecken put his linkage onto a typical Reuleaux frame and assigned to it the model number 2201. There is also a paper of Hoecken on dwell mechanisms [8], but it does not describe his model shown in Fig. 6. Before manufacturing the model \u201csix-link dwell mechanism\u201d, Hoecken made a technical drawing of it on February 11, 1932 and named it \u201cRastgetriebe\u201d, Fig. 7. The drawing shows details of the links, its dimensions and even a counterweight on the opposite end of the rocker for static balancing. The dimensions of Hoecken\u00b4s six-link dwell mechanism are as follows: n0 = B0C0 \u2248 124.0 A0C0 \u2248 215.0 n1 = A0B0 \u2248 127.0 n2 = A0A = 61.3 n3 = AB = 131.9 n4 = B0B = 94.2 n5 = CP3k = 83.9 n6 = C0C = 139.5 AP3k = 84.8 \u2220P3kAB = 24.45 deg The author again used the commercial program \u201cMathcad 11\u201d to make a kinematic analysis of Hoecken\u00b4s six-link dwell mechanism, the results are presented in Fig. 8. The approximate dwell of the driven link 6 (C0C, Fig. 7) is realized within the range 120 deg \u2264 \u03c6 \u2264 180 deg of the driving link 2 (A0A, Fig. 7). Within this range the angle \u03c66 of the driven link falls from 214 deg to 213 deg only, the angular velocity of the driven link shows an absolutely largest deviation from zero of 0.022 rad/s, the one of the angular acceleration amounts to 0.018 rad/s2. Redtenbacher\u00b4s gear train In 1926 Hoecken visited the Polytechnic School (Technische Hochschule, TH) in Karlsruhe. He found among other things the model of a gear train with seven gearwheels. The model belonged to the Redtenbacher model collection in Karlsruhe, but nobody could explain to Hoecken the purpose of this model [9]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.17-1.png",
        "caption": "FIGURE 6.17",
        "texts": [
            " In this case for a cylindrical bush the torsional stiffness of the bush would be zero to allow rotation about the axis, or could have a value associated with the friction in the joint. In this model the connection of the anti-roll bars to the suspension system is not modelled in detail, rather each antiroll bar part is connected to the suspension using an inplane joint primitive that allows the vertical motion of the suspension to be transferred to the anti-roll bars and hence produce a relative twisting motion between the two sides. A more detailed approach, shown in Figure 6.17, involves including the drop links to connect each side of the anti-roll bar to the suspension systems. The drop link is connected to the anti-roll bar by a universal joint and is connected to the suspension arm by a spherical joint. This is similar to the modelling of a tie rod as discussed in Chapter 4 where the universal joint is used to constrain the spin of the link about an axis running along its length, this DOF having no influence on the overall behaviour of the model. The stiffness, KT, of the torsional spring can be found directly from fundamental torsion theory for the twisting of bars with a hollow or solid circular cross-section",
            " Assuming here a solid circular bar and units that are consistent with the examples that support this text we have Modelling the anti-roll bars using joint primitives. REV, revolute joint. Modelling the anti-roll bars using drop links. REV, revolute joint. L GJKT = \u00f06:3\u00de where G is the shear modulus of the anti-roll bar material (N/mm2) J is the polar second moment of area (mm4) L is the length of the anti-roll bar (mm) Note that the length L used in Eqn (6.3) is the length of the bar subject to twisting. For the configuration shown in Figure 6.17 this is the transverse length of the anti-roll bar across the vehicle and does not include the fore-aft lengths of the system that connect to the drop links. These lengths of the bar provide the lever arms to twist the transverse section of bar and are subject to bending rather than torsion. An externally solved FE model could be used to give an equivalent torsional stiffness for a simplified representation such as this. Given that bending or flexing of the roll bar may have an influence, the next modelling refinement of the anti-roll bar system uses FE beams, of the type described in Chapter 3, to interconnect a series of rigid bodies with lumped masses distributed along the length of the bar"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure7.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure7.7-1.png",
        "caption": "Fig. 7.7 (a) A sample three-dimensional geometry. (b) An extracted slice as a two-dimensional shape",
        "texts": [
            " Alternatively, designers can get the CAD model through geometric scanning or medical imagining techniques which may reduce the time duration of product design. Then, the geometry file can be transformed and stored in the \u201cstereo lithography\u201d (STL) format which can allow designers using CAD software to position, scale, and even optimize the STL model by checking and fixing shape features such as holes, overlaps, and offsets. CAD software can then transform the geometry into multiple slices of two-dimensional cross sections (Fig. 7.7). During this step, designers can configure the slicing procedures with the machine parameters such as the layer thickness, printing position, object position, and supporting mode. Designers should also select the printing materials. If necessary, support materials 198 7 Medical Imaging and Reverse Engineering can be added underneath the product shape as temporary supports during the fabrication process (Fig. 7.8). After printing the material as defined by the STL model (and the support material), post-processing should then be conducted, including taking the object out of the printing board, removing the support material, cleaning the printed object, further curing (e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002624_jae-150151-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002624_jae-150151-Figure6-1.png",
        "caption": "Fig. 6. (a) When \u03b1 = 10%, magnetic field lines for the generator of rated-load; (b) When \u03b1 = 10%, electromagnetic torque for the generator of rated-load.",
        "texts": [
            " The initial negative current electrical angle recorded by \u03b6r refers to phase angle difference between the instantaneous value of the negative current in phase-A and the instantaneous value of the positive current in phase-A. Define the current asymmetric degree as \u03b1 = I1s\u2212/I1s+. Under the circumstance that the positive sequence network is at rated-load, add the negative component to the negative sequence network and the initial negative electrical angle is zero, when \u03b1 = 10%, the magnetic field lines for the generator of rated-load is shown in Fig. 6(a), and the electromagnetic torque for the generator of rated-load is shown in Fig. 6(b). The simulation result is similar with the analytic algorithm that when the generator operates at balanced state, on the condition of no-load, the ripple of the electromagnetic torque is 0.827\u2030 of the constant electromagnetic torque at rated-load, and on the condition of rated-load, the ripple of the electromagnetic torque is 2.7\u2030 of the constant electromagnetic torque at rated-load. The value of the constant electromagnetic torque and the peak value of the second harmonic electromagnetic torque are the arithmetic average about the initial negative current electrical angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001375_ecai.2014.7090193-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001375_ecai.2014.7090193-Figure1-1.png",
        "caption": "Figure 1. Coordinate system",
        "texts": [
            " This approach is valid for actuator/surface failures, which cause to the different control effectiveness factors (for example, partial loss of a control surface - when a part of the control surface breaks off - deformation of the control surface, etc.). II. UUV MATHEMATICAL MODEL UUVs move in 6 degrees of freedom (6 DOF) thus six independent coordinates are necessary to determine the position and orientation of a rigid body. The first three coordinates and their time derivatives are of translational motion along the x, y and z-axes, while the last three coordinates (\u03a6, \u03b8, \u03c8) and time derivatives are used to describe orientation and rotational motion (See Fig.1). In Fig. 1; u, v, w are the velocity of surge, sway and heave motion and p, q and r are angular velocity of roll, pitch and yaw motion, respectively. X, Y, Z and K, M, N represent the forces and moments acting on the UUV, with respect to x, y, and z axes. The Euler angles that defines UUV\u2019s orientation relative to the Earth-fixed coordinate frame\u2019s x, y, and z axes are denoted by \u03a6, \u03b8, \u03c8. Body-fixed frame's position is relative to Earth-fixed frame is represented by ro. The 6 DOF equations of motion for the model are given in Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000887_gtindia2014-8186-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000887_gtindia2014-8186-Figure3-1.png",
        "caption": "Figure 3. Illustration of SFD bearing housing.",
        "texts": [
            " Oil is supplied through 2 internal ports (jets) projecting to the rolling element bearings on the rotating high speed shaft. The oil leakage from the shaft ends are prevented by a non-contact, labyrinth seal. The drained oil is drawn out (by gravity) through the port. A rotor with two disks simulating the required weight will be supported on high precision angular contact ball bearings. SFD bearing housing has provision to induce intentional misalignment up to a maximum of 0.15mm. The design of this housing is shown in Figure 3. Bearing housing is made in two half for easy assembly / disassembly of shaft with inertia discs. Bottom frame (item no.-1) is mounted on the cast iron bed plate. Bottom linear slides (Itemno.-2) are assembled on the bottom frame. Bottom bearing housing(item no.-4) is mounted over the linear slides, two locking plates (item no.-3 & 9) are provided on the both sides of the bearing housing to lock the bearing housing in place after radial misalignment is induced. For inducing the radial misalignment, a separate arrangement (item no"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000989_s106836661401005x-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000989_s106836661401005x-Figure1-1.png",
        "caption": "Fig. 1. Calculation scheme of sliding bearing: (1) shaft; (2) polymer bushing; (3) cage; (4) axis of loading of bearing.",
        "texts": [
            " Because of this, during the modeling of the thermal process in the bearing, the polymer bushing and the metal cage can be considered plane members, while the shaft can be considered a three dimensional mem ber. Thus, the temperature field in the sliding bearing can be represented as a superposition of two and three dimensional fields. When the thermal diagnos tics of friction uses this mathematical model of the thermal process, it is sufficient to assign the excess temperature in the same manner as was applied in the plane formulation of the problem [1, 2]. SIMPLIFIED THREE DIMENSIONAL MODEL Let us consider a sliding bearing shown schemati cally in Fig. 1. Under the given assumptions, the tem perature distribution in the bushing with the cage is described by the following two dimensional heat conduction equation with the coefficients C(T) ( , , )T r t\u03d5 Keywords: sliding bearing, thermal diagnostics of friction, frictional heat generation, temperature, thermal process, model, inverse boundary problem, simulation of error, regularization, computational experiment DOI: 10.3103/S106836661401005X JOURNAL OF FRICTION AND WEAR Vol. 35 No. 1 2014 THERMAL DIAGNOSTICS OF FRICTION IN SLIDING BEARINGS 49 and \u03bb(T), which have discontinuities at the bushing\u2013 cage interface at (1) The temperature field in the shaft is described by the following three dimensional heat conduction equation with the convective term, which considers the movement of the shaft: (2) In the contact zone the following conjugation condi tions on the lateral surface of the shaft are written: (3) (4) where On the lateral surface of the shaft that is outside of the contact where D is the total area of the lateral surface of the shaft, heat transfer to the environment occurs in accordance with Newton\u2019s law as follows: (5) :2r R= ( ) 2 1 3 1 1( ) ( ) ( ) , , , 0 ",
            ", the following condition is fulfilled [11]: (43) where is the dispersion of the function Similar to the direct problem, the problem for the increments and the conjugate problem are solved by the finite difference method. COMPUTATIONAL EXPERIMENTS Let us present the results of the computational experiments aimed at investigating the solution of the inverse boundary problem for restoring the specific power of frictional heat generation The com putations were carried out for the sliding bearing shown in Fig. 1 with the following geometrical dimen sions: l1 = 0.02, l2 = 0.005, l3 = 0.026, l4 = 0.020, l5 = 0.025, L = 0.096, = 0.012, Rch3 = 0.006, R1 = 0.0125, R2 = 0.0165, and R3 = 0.032 m, as well as \u03d50 = 30\u00b0. The initial uniform temperature dis tribution in the friction unit T0 was equal to the ambi ent temperature Tam = 20\u00b0C. The thermal characteris tics of the materials of the shaft and bearing were taken to be the same as those used in the plane problem [1, 2]. The model function of the specific power of heat generation was assigned as follows: Q(\u03d5, t) = where The angular velocity of the shaft is (s\u20131), which corresponds to 30 rpm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.30-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.30-1.png",
        "caption": "Figure 9.30 Sensorless start-up of a rotating PMSM. The rotation angle, \u03c8PM, and ePM are described at time instants t1 and t2.",
        "texts": [
            "29, it is easy to identify the rotor position of a spinning rotor by applying two short successive short circuits, selecting zero voltage vectors, and then measuring the short-circuit current produced by the original back emf. The spinning rotor induces a voltage that is 90 electrical degrees behind \u03c8PM, and the short-circuit current starts growing in the direction of the induced voltage vector. Using two successive pulses, rotor speed and \u03c8PM position can be found, and the drive can be started. See Figure 9.30. When a short circuit is triggered by a zero voltage vector at time t1, the following current develops in the direction of ePM1 (ePM1=\u03c8PM\u03c9re j\u03c0/2). 1 1 j\u03c0=2dtePM1dt (9.71)i1 \u222b L \u222b L\u03c8PM1\u03c9re t t If test time t is sufficiently short, the vector direction of i1 for an arbitrary angle \u03b81 can be obtained. i1 iej\u03b81 (9.72) Following the second short circuit, i2 can be determined. i2 iej\u03d12 (9.73) The rotor electrical angular velocity is \u03d12 \u03d11\u03c9r (9.74) t t2 1 Rotor position angle is as follows. \u03c0 \u0394t\u03c9r (9"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002952_gt2016-56518-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002952_gt2016-56518-Figure1-1.png",
        "caption": "Figure 1. Turbine structure (a), magnification of the fir-tree root and groove (b) and initial fir-tree root shape in optimization(c), the red dots represents the defined control points, whose horizontal coordinates are selected as design variables in optimization procedure to modify the fir-tree root shape.",
        "texts": [
            " In this paper, Spline curve technique is used to model the fir-tree root, and in the optimization procedure, since the centrifugal load corresponds to the 100% rated speed, the maximum strain falls into the elastic regime, the linear elastic model is used and the objective function is minimizing of the maximum von-Mises stress instead of plastic strain, the horizontal positions of several specified control points in the symmetrical plane are defined as design variables, the optimization constraints are extracted from industrial experience. The Multi-island genetic algorithm (GA) are perform to seek the optimal fir-tree root and groove configurations. The typical turbine structure with fir-tree root and groove is depicted in Fig.1a while the Fig.1b shows the magnification of the fir-tree root and groove, the stress distributions in which is the key issue in this paper. To model the fir-tree root and groove, we use an in-house script to establish the finite model and characterize the shape with the Spline curve technology, then the generated input file is exported to the commercial package ABAQUS to perform the finite element analysis to evaluate the stress distributions and stress concentrations. To ensure the efficiency of the optimization approach, the finite element model used in each optimization loop is based on 2D assumption instead of 3D characterization. In the fir-tree 2D plane strain model, the blade root with intermediate platform is constructed shown in Fig.1c, to save the computational time and focus on the root morphology design, the eccentric influence induced by the intermediate platform in 3D model is not considered here and the platform is characterized by straight lines. When characterizing the fir-tree root, spline curves are needed and several control points are located at one side of the blade root configuration as shown in Fig.1c while the fir-tree root 2 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89525/ on 03/01/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use has a symmetrical shape according to the y axis. The horizontal coordinates of these control points are defined as the design variables. The control points are located at the inflection points which dominate the profile and feature of the fir-tree root. We find that 13 is the minimum control point number to accurately characterize the fir-tree root",
            " When the finite element analysis is completed, several indices are retrieved from the results, such as maximum von-Mises stress for each partition region and whole model, radial displacement for central point in the bottom surface of fir-tree root, variance of stresses in root and rotor groove region. These indices will be used in the optimization approach as design variable, optimization constraints and objective function, respectively. The optimization formula can be stated as follow:     min. max . 1,2, , N low up i i i up up s t X x X i D D w W           (1) where max( ) represents the maximum von Mises stress in the root and disc region; ix is the horizontal coordinates of ith control point to construct fir-tree root as shown in Fig.1c, N is the total number of control points; low iX and up iX are the corresponding lower and upper bounds, respectively, which can be obtained from industrial case. D is the calculated variance of stresses in the root or disc; upD is the upper bound of the root or disc; w is the radial displacement of the reference point (central point in the bottom surface of fir-tree root) and upW is the upper bound. The stresses variance can be defined as:     2 1, 2,3i iD i    (2) where i is the max stress of the ith part of the root or disc and i is the mean stress of the root or disc"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003523_cefc.2016.7815987-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003523_cefc.2016.7815987-Figure1-1.png",
        "caption": "Fig. 1. Topology and flux distribution without load. (a) Machine topology, winding connection and DC current injection direction. (b) PM excited only. (c) Both PM and 4 A DC current excited.",
        "texts": [
            " INTRODUCTION The stator permanent magnet (PM) machines are rapidly developed for their features of easy heat dissipation of PMs [1]. Besides, in vehicle application which requires wide constant power speed range, the hybrid excitation PM machines (HE-PMMs) have be gaining attention. However, the existing HE-PMMs often require a special field winding fed with DC current, and the field winding brings a space conflict between the stator PM and armature winding, also, it will increase the machine cost. In this paper, a novel HE-VPM is proposed. As shown in Fig. 1 (a), the consequent PMs are embedded in the stator teeth, helping for heat dissipation, and reduce the amount of PM; besides, the salient rotor without PMs and winding ensures the robustness; furthermore, the non-overlapping concentrated winding can reduce the end-winding length and is help for torque density. The winding arrangement and the biased DC current direction are shown in Fig. 1. The number of stator/rotor slots are 12/10, and the armature winding pole pair is 4. The corresponding drive circuit can be referenced in [3]. Fig. 1 (b) and (c) show the flux distribution under no load condition but with different excitation. Fig. 2 (a) compares the simulated phase back-EMF waveforms with different excitation conditions. As shown, the phase angle of back-EMF produced by the DC current aligns with that by PM. Besides, it can be found that the back-EMF can be adjusted flexibly by the DC current. Therefore, the constant power speed range enlarges. Since the phase current contains DC field component and AC armature component, there will be an optimum current combination for maximum torque"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001391_dscc2014-6301-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001391_dscc2014-6301-Figure1-1.png",
        "caption": "Figure 1. Bicycle model schematic",
        "texts": [
            " To investigate the interaction between this model and active steering interventions, large interventions are created using the obstacle avoidance controller proposed by Erlien et al. [5]. The paper then demonstrates, through simulation and experiments, that using the virtual wheel to create supportive handwheel torque reduces the discrepancy between the driver command and the active steering system in simulation and experiments. The vehicle model used in this work is a constant speed, planar bicycle model. The vehicle\u2019s motion is described by two states: sideslip \u03b2 and yaw rate r. These are defined in Figure 1 and have the following equations of motion: \u03b2\u0307 = Fyf +Fyr mUx \u2212 r r\u0307 = aFyf\u2212bFyr Izz (1) where Fy[f,r] is the lateral tire force of the [front, rear] axle, m is the vehicle mass, Ux is the longitudinal velocity in the body fixed frame, Izz is the yaw inertia, and a and b are the distances from the center of gravity to the front and rear axles, respectively. The simplified Fiala tire model introduced by Pacejka in [8] gives a useful approximation of the non-linear relationship be- tween Fy[f,r] and the tire slip angles \u03b1[f,r]: Fy =  \u2212C\u03b1 tan\u03b1+ C2 \u03b1 3\u00b5Fz | tan\u03b1| tan\u03b1 ",
            " \u2212 C3 \u03b1 27\u00b52F2 z tan3 \u03b1, |\u03b1|< tan\u22121 ( 3\u00b5Fz C\u03b1 ) \u2212\u00b5Fzsgn \u03b1, otherwise (2) = ftire (\u03b1) (3) where \u00b5 is the surface coefficient of friction, Fz is the normal load, and C\u03b1 is the tire cornering stiffness. The tire slip angles, \u03b1f and \u03b1r, can be derived from the kinematics of the vehicle as: \u03b1f = tan\u22121 ( \u03b2+ ar Ux ) \u2212\u03b4 (4) \u03b1r = tan\u22121 ( \u03b2\u2212 br Ux ) (5) where \u03b4 is the steer angle. The vehicle\u2019s position is specified relative to a reference line using three states: heading deviation \u2206\u03c8, lateral deviation e, and distance along the path s as defined in Figure 1. The equations of motion of these states can be written as: \u2206\u0307\u03c8 = r (6) e\u0307 = Ux sin(\u2206\u03c8)+Uy cos(\u2206\u03c8) (7) s\u0307 = Ux cos(\u2206\u03c8)\u2212Uy sin(\u2206\u03c8) (8) Figure 2 illustrates the mechanical coupling of the handwheel and the roadwheels in a conventional steering system. In steer-by-wire vehicles, the handwheel is mechanically decoupled from the roadwheels. A force feedback (FFB) steering system can be used in tandem with a steering feedback model to create realistic steering feedback for the driver on a steer-by-wire vehicle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001174_0954405414554016-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001174_0954405414554016-Figure3-1.png",
        "caption": "Figure 3. Anti-backlash gear transmission.",
        "texts": [],
        "surrounding_texts": [
            "Hertz contact, dynamic characteristics, anti-backlash gear, angle-contact ball bearing Date received: 26 November 2013; accepted: 9 September 2014"
        ]
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.70-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.70-1.png",
        "caption": "FIGURE 6.70",
        "texts": [
            " As discussed earlier the incorporation of microprocessor control systems in a vehicle may involve the use of a simulation method that involves: 1. the use of MBS software where the user must invest in the modelling of the control systems. 2. the use of software such as MATLAB/Simulink where the user must invest in the implementation of a vehicle model or, 3. a co-simulation involving parallel operation of the MBS and control simulation software. In this example the author (Wenzel et al., 2003)2 has chosen the second of the above options and a vehicle model (Figure 6.70) is developed from first principles and implemented in Simulink. The model developed here is based on the same data used for this case study with 3 DOF: the longitudinal direction x, the lateral direction y and the yaw around the vertical axis z. The vehicle parameters used in the following model include: 2Wenzel et al., 20 Jaguar Cars Ltd, C University, Coven vx 03, describes prel oventry, UK and try, UK. It forms Longitudinal velocity (m/s) vy Lateral velocity (m/s) vcog Centre of gravity velocity (m/s) ax Longitudinal acceleration (m/s2) ay Lateral acceleration (m/s2) G Torque around z-axis (Nm) d Steer angle (rad) b Side slip angle (rad) aij Wheel slip angles (rad) Yaw rate (rad/s) Fzij Vertical forces on each wheel (N) Ij Position: i\u00bc front(f)/rear(r), j\u00bc left(l)/right(r) Note that steer angle d and the velocity of the vehicle\u2019s centre of gravity vcog are specified as model inputs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001822_jae-150004-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001822_jae-150004-Figure2-1.png",
        "caption": "Fig. 2. The schematic of search coil for eccentricity detection.",
        "texts": [
            " In the process of obtaining air-gap magnetic field, it could be known that the characteristic frequencies (\u03c9 + (Z2 \u00b1 1) \u03c9r) are produced due to the interaction of dynamic eccentricity and rotor slotting. Based on this, these frequencies are denoted by IDERS in this paper. Therefore, IDERS can be as diagnostic signal in detecting dynamic eccentricity and mixed eccentricity aimed at the rotors slotting of induction motors. The schematic of search coil used for measuring air-gap magnetic fields is shown in Fig. 2. In practical application, the search coil is open circuit. There is only induced voltage. The current of search coil is equal to zero. Therefore, the search coil would not create abundant magnetic field harmonics to disturb the original magnetic field. The following assumptions are made. 1) There are no leakage magnetic fields in the stator slots. 2) All flux lines pass through the stator teeth. 3) The angular of search coil is zero in the stator frame of reference. Then the magnetic flux linkage in the search coil could be expressed as \u03c8 = N ltrav 2 \u222b 2\u03c0 Z1 0 B\u03b4(t, \u03d5)d\u03d5 (12) where N is the turns of the search coil, lt is the stator stack length, rav is average radius of air gap, B\u03b4 (t, \u03d5) is the air-gap magnetic field"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure8.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure8.9-1.png",
        "caption": "Fig. 8.9 Side-view illustration of the melt pool shape and liquid motion",
        "texts": [
            " A laser can be used as a highspeed, high-quality welding tool. The advantages in laser welding, for example, include an accessible spacing between the sample surface and the laser source, enabling real-time inspection of the process, simplifying the quality control process, and relatively lower costs. Compared to the traditional approaches such as tailored blank welding for the car industry, laser welding has great demand in areas requiring heat-sensitive components such as heart pacemakers. The basic operation of laser welding is illustrated in Fig. 8.9. Conduction-limited welding occurs when the power density is insufficient to cause boiling \u2013 and therefore generate a keyhole \u2013 at the given welding speed. The weld pool has strong stirring forces driven by Marangoni-type forces resulting from the variation in surface tension with temperature. There are two principle areas of interest in the mechanism of keyhole welding. The first is the flow structure since this directly affects the wave formation on the weld pool and hence the final frozen weld bead geometry"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002584_ijamechs.2015.074786-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002584_ijamechs.2015.074786-Figure4-1.png",
        "caption": "Figure 4 UVMS for simulation",
        "texts": [
            " In view of this situation we will compare the performances in simulation studies shown in the subsequent section. 4 Numerical simulations In order to compare the performances of the proposed robust controller and the existing controller, we performed numerical simulations. The proposed controller is the robust one with the fixed compensator expressed as (25) and (36). On the other hand, the existing controller is the robust one without a fixed compensator designed in Taira et al. (2011). The UVMS simulated here was an underwater vehicle (robot) with a two-link manipulator, as shown in Figure 4. The values of its system parameters (excepting thrusters\u2019 parameters) were the same as those used in Taira et al. (2010, 2012). In this figure, only the values of the main parameters are shown. The parameters of the thrusters were given by A = (1/T)I3, B = (235,602/T)I3, and 30.00111 ,K I= \u00d7 where 3 3 3I R \u00d7\u2208 is an identity matrix, and T is a positive constant. These values were determined so that the steady state responses could be roughly the same as those of the experimental results in Kim and Chung (2006), whereas the transient responses are changed by T"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000430_978-3-319-15684-2_4-Figure4.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000430_978-3-319-15684-2_4-Figure4.1-1.png",
        "caption": "Fig. 4.1 Kinematic model of a biped robot",
        "texts": [
            "), Advances in Mechanical Engineering, Lecture Notes in Mechanical Engineering, DOI 10.1007/978-3-319-15684-2_4 23 walking robots lean only by one foot on the ground for an appreciable period of time [3]. For a biped robot, the support area is small. Because it is passively unstable and non-linear it is not easy to design a walking controller. The control system should provide processing of information about the area [7, 11], making decisions about the movement, and control over the implementation. Figure 4.1 shows the structure of the biped and coordinates used to describe the configuration of the system. In [4, 5] we have obtained the equations of spatial movement, investigated the possibility of automatic control for stable and constant biped walking, and defined the desired time and place for touchdown at the end of the step and at the beginning of the next one. This paper proposes a method of biped walking control in different modes: walking in the up or down direction, walking up stairs or down stairs, and rotation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003350_cefc.2016.7815913-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003350_cefc.2016.7815913-Figure1-1.png",
        "caption": "Fig. 1. The structure and flux contour linesexcited by magnets of the proposed machine.",
        "texts": [],
        "surrounding_texts": [
            "The structure of this proposed machine is shown in Fig .1. It is seen that the consequent pole in the rotor is employed, and the Halbach array PMs in the stator opening is used to guide the flux through airgap into the stator as shown in Fig. 2. As shown in Figs. 1 and 2, only the S-pole magnets produce the main flux, and N-pole magnets produce the leakage flux to reduce main flux. In the proposed PMV machine, almost all the magnets produce the main flux. Hence, the significant improvement on the magnet flux can be obtained with similar magnet usage as shown in Fig 3."
        ]
    },
    {
        "image_filename": "designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.12-1.png",
        "caption": "FIGURE 9.12",
        "texts": [
            "3 BUCkLING Of INCLINED CIRCULAR CYLINDER-IN-CYLINDER The postbuckling configurations, obtained by the DQM with small disturbance, are shown in Fig. 9.11. There are two plots in the figure representing the same postbuckling configuration but viewing at different directions. It is seen that the buckling mode is lateral and not helical. At P = 1200 N, the maximum rotation is only 0.9479 radians, i.e., \u03b8 \u03c0= 0.30max for the lateral buckling shown in Fig. 9.11. The postbuckling configurations, obtained by the DQM with large disturbance, are shown in Fig. 9.12. There are also two plots in the figure representing the same postbuckling configuration but viewing at different directions. It is seen that the buckling mode is helical and not lateral. At P = 1200 N, the maximum rotation is 13.532 radians, i.e., \u03b8 \u03c0= 4.31max for helical buckling shown in Fig. 9.12. umax=0.30\u03c0 umax=4.31\u03c0 P\u2212umax Curves with Small and Large Initial Disturbances Postbuckling Configuration (Small Disturbance, P = 1200 N, a = 45\u00b0) 188 CHAPTER 9 GEOMETRIC NONLINEAR ANALYSIS The results obtained by the DQM, i.e., Figs. 9.10\u20139.12, are similar to the data obtained by the discrete singular convolution (DSC) [3]; thus, the formulations are verified each other. It should be mentioned that large disturbance should be used in order to catch the helical buckling mode [3], otherwise only lateral buckling mode can be captured by various numerical method, such as the DQM, DQEM, FEM, LaDQM, and DSC",
            " Comparison of Fig. 9.15 with Fig. 9.13 shows that the variation of b is quite different due to different buckling configurations shown in Figs. 9.11 and 9.12. Figure 9.16 shows the distribution of the contact force per unit length (Wn) at three applied loads. The results are obtained by the DQM with large disturbance. It is seen that ( ) \u2265W 0n everywhere. From Figs. 9.15 and 9.16, the assumptions in deriving the governing equation are valid. Therefore, the results shown in Fig. 9.10 ( \u2265P N1200 ) and Fig. 9.12 are also valid solutions. sin(b)\u2248b Wn\u22650 sin(b)\u2248b Wn\u22650 P\u22651200 N Variations of Wn at Various Applied Axial Loads (Small Disturbance) Variations of b Obtained by the DQM with Large Disturbance 190 CHAPTER 9 GEOMETRIC NONLINEAR ANALYSIS The results obtained by the DQM are more or less similar to the data obtained by the DSC; thus, Figs. 9.6\u20139.16 are almost the same as the corresponding figures presented in [3,5,6]. For helical buckling analysis, especially the rod is long, the DQEM or LaDQM is recommended since a large number of grid points are needed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000968_20140824-6-za-1003.01643-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000968_20140824-6-za-1003.01643-Figure4-1.png",
        "caption": "Fig. 4. Flight corridors of two UAVs and their safety circles \u2013 first scenario Finally, the third possible scenario is shown in Fig. 4c, in which it is easy to verify that ji DOCDOC < . Here, it is clear that iUAV should have higher priority over jUAV . As a result, iUAV continues moving toward the objective circle while jUAV starts a collision avoidance maneuver. Similar reasoning can be made to come to the same conclusions where jUAV approaches iUAV from the left side. By",
        "texts": [
            " Also, addition of the conservative second term i.e. lR2 accounts for the fact that one of the UAVs may already be loitering to avoid a possible collision. An important observation is that if the safety circle of jUAV does not overlap the flight corridor of iUAV at a given moment, there will be no potential risk of collision for iUAV with jUAV , at that configuration. Regarding the relative configuration of any two UAVs, there are 3 possible scenarios to be considered. In the first scenario shown in Fig. 4a, safety circle of jUAV is tangent to the flight corridor of iUAV , but safety circle of iUAV does not intersect flight corridor of jUAV . Here, it is clear that jUAV should continue moving toward the objective circle while iUAV should start a collision avoidance maneuver. Let us show the distance between iUAV and the center of the objective circle by iDOC . Similarly, the distance between jUAV and the center of the objective circle is denoted by jDOC . For the configuration shown in Fig. 4a, it is easy to verify that ji DOCDOC < . A second possible scenario is shown in Fig. 4b. In this scenario, if any of the UAVs moves, the flight corridor of the other UAV will be blocked. In this specific configuration where ji DOCDOC = , the UAV with greater heading angle will have priority over the UAV with smaller heading angle ( deg3600 <\u2264 \u03b8 ). Thus, higher priority is given to jUAV and lower priority is given to iUAV . Thus, jUAV keeps moving toward the objective circle while iUAV starts a collision avoidance maneuver. construction, the above collision avoidance scheme does not allow the safety circles of two UAVs to simultaneously block each others\u2019 flight corridor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002042_arso.2015.7428209-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002042_arso.2015.7428209-Figure4-1.png",
        "caption": "Fig. 4. Definition of the workspaces in which the robot is controlled in the locomotion and manipulation modes. Note that the workspace is defined for the manipulator arm with respect to the vehicle frame Fb, and not the world frame F0. The velocity is generated by the virtual spring between the master manipulator (gray) and the slave manipulator (black). The intuitive interpretation of the virtual spring is illustrated by the spring between the master manipulator and the vehicle.",
        "texts": [
            " When in locomotion mode we allow only for motion of the vehicle which is given by[ vs \u03c6s ] = [ \u2212kv 0 0 \u2212k\u03c6 ] [ d1 d2 ] (6) where kv and k\u03c6 are proportionality constants; vs and \u03c6s are the velocity and the heading angle of the vehicle in the body frame; and d1 and d2 are defined by the position of the haptic device, as shown in Figure 3, i.e., the distances from the master\u2019s tip position to the limit area that is used to define the manipulation mode. In this section we describe the control scheme, first presented in [8], which introduces artificial forces between the end-effector and the base. First we find the manipulator workspace WM with respect to the vehicle frame Fb. We define the workspace for position control as a workspace WP , somewhat smaller than the manipulator workspace WM , as illustrated in Figure 4. Whenever the manipulator is inside this workspace, position control is applied. This is equivalent to the manipulation mode in the previous sections. If the master manipulator is outside the workspace WP , velocity control is applied. In this case the slave manipulator remains fixed at the limit of the workspace, while the vehicle velocity is so that the vehicle follows the master end-effector with a mass-spring-damper characteristics. Denote by x\u0304s the position of the end effector projected into the position workspace WP , as illustrated in Figure 4. Then the slave position with respect to this projected position is given by \u0394 = xs\u2212x\u0304s and we will let the vehicle be governed by F = \u0394\u0308 + d\u0394\u0307 + k\u0394. (7) The following references will give the above characteristics: \u2022 Manipulator arm reference: V B 0e,r = V B 0e,d \u2212 1 db F, (8) \u2022 Vehicle reference: V B 0b,r = 1 db F. (9) This control law is to be interpreted in the following way: The desired end-effector velocity in the inertial space is given by V B 0e,d. The manipulator reference is obtained by the Adjoint map Adg ([5]) and subtracting the vehicle motion V B 0b,r, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000047_ijhvs.2020.104400-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000047_ijhvs.2020.104400-Figure3-1.png",
        "caption": "Figure 3 The wheel/rail contact zone",
        "texts": [
            " The wheelset\u2019s dynamic behaviour is significantly affected by creep forces and creep moments which arise between the wheels and the rails. Among various wheel-rail rolling contact theories (Garg and Dukkipati, 1984), Kalker\u2019s linear creep theory, which has a negligible effect of variation of the normal reaction on the value of the creep coefficients, is applied in this work. Creep forces and creep moments are remarkably influenced by the area of contact and contact stresses between the wheel and rail and generally defined as functions of creepages on the elliptical contact surface (Figure 3). The definitions of creepages, 1\u03be , ( 2\u03be , sp\u03be ) are the ratios of the difference between a wheel\u2019s longitudinal (transverse, spin) speed and a rail\u2019s longitudinal (transverse, spin) speed at the contact point, and the nominal speed V . Based on these definitions, the creepages of the left wheel are 1 1 3 0 ( ) /N P L aV u y V r \u03bb\u03be = \u22c5 = \u2212 \u2212v l (51) 0 5 2 2( ) / (1 )N P L r u V a V \u03bb\u03be \u03c8= \u22c5 = + \u2212v l (52) 3 3 0 1( ) / .N D spL V u V r \u03bb\u03be = \u22c5 = \u2212\u03c9 l (53) In this case, the creep force LF applied to the left wheel is 33 1 1 11 2 12 2( )L L L spLf f f\u03be \u03be \u03be= \u2212 + \u2212 \u2212F l l (54) and the creep moment LM is 12 2 22 3( )L L spLf f\u03be \u03be= \u2212M l (55) Note that 11f , 12f , 22f , and 33f denote lateral, lateral/spin, spin, and longitudinal creep force coefficients, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000859_1.4027130-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000859_1.4027130-Figure1-1.png",
        "caption": "Fig. 1 Single-row trochoidal gear",
        "texts": [
            " Therefore, it is clear that using a double-row trochoidal gear is effective for reducing the transmission errors of trochoidal gears. [DOI: 10.1115/1.4027130] Keywords: gear, trochoidal gear, roller gear, cam gear, contact ratio, transmission error, multibody analysis In recent years, in order to develop a highly accurate and highperformance mechanical system, a more precise mechanism for positioning the gear reducers has been required. To meet this need, a single-row trochoidal gear has been developed, as shown in Fig. 1 [1\u20133]. A pair of commercial trochoidal gears consists of a single-row roller gear and single-row cam gear with a trochoidal tooth profile. Since both ends of the rollers on the roller gear are supported by needle roller bearings, the rollers can spin on their axis. As an important indicator of the operating performance of gears, transmission error has been widely used. The transmission error (TE) is defined as [4] TE \u00bc hout chin (1) where hin is the input rotational angle, hout is the output rotational angle and c is the reduction ratio"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure9-1.png",
        "caption": "Figure 9 Deadlock owing to friction (see online version for colours)",
        "texts": [
            " However, the previous manipulator still has several disadvantages: P1 an asymmetrical range of movement P2 deadlock owing to friction. As shown in Figure 8, the range of movement in the horizontal direction is adequate to search for survivors. However, the turning range is restricted by the rail in the vertical direction. The range of movement depends on the extension of the rail. Problem P1 arises from the fact that the centre of rotation of the joint is not coincident with the centre of the rail. Figure 9 shows P2. In the case of a tight curve, the pushing force is converted to frictional force. If \u03bc \u2265 (1/tan \u03b8), the driven manipulator cannot move forward. To solve P1, we propose a new manipulator with a rail at the centre of rotation of the joints; to solve P2, we employ a pulling mechanism. By employing both pushing and pulling forces, deadlocks can be reduced. To solve the problems of the previous manipulator, we developed a new manipulator for the duplex mechanism, which is shown in Figures 10 and 11"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003298_j.ifacol.2016.10.195-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003298_j.ifacol.2016.10.195-Figure1-1.png",
        "caption": "Fig. 1. Control of pendulum using flywheel",
        "texts": [
            " (2009)] dx1 d\u03c4 = x2 dx2 d\u03c4 = \u2212 sinx1 \u2212 \u03b3x2 + ax3 \u2212 bu, dx3 d\u03c4 = sinx1 + \u03b3x2 \u2212 \u03b4ax3 + \u03b4bu, |u| \u2264 1. The aim of the control parameter u is to lead out the pendulum to one of positions shown below: 1. x1 = 0, x2 = 0, x3 = 0 2. x1 = \u03c0, x2 = 0, x3 = 0 Let us provide some results of numerical computations obtained for this model. All computations are performed for the following values of variables: a = 0.05, b = 1, \u03b3 = 1, \u03b4 = 2. The time interval under consideration is t = [0; 3] with the time partition \u2206t = 0.01. The partition of the control parameter u is \u2206u = 0.1. Fig. 1. - Fig. 6. show evolution of reachability sets for time moments t = 0, t = 0.75, t = 1.5, t = 2.25, t = 3, respectively. We present the set here as the 2-dimensional projection to provide better understanding of its geometry. Calculation of trajectories for the system is implemented for two different initial states of x0. Trajectory X1 from x0 = (\u22123.477, 3.293,\u22124.814) to xf = (0, 0, 0) is shown on Fig. 7. The control vector u1 which corresponds to this trajectory is shown on Fig. 8. Trajectory X2 from x0 = (\u22120"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002340_indicon.2015.7443810-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002340_indicon.2015.7443810-Figure2-1.png",
        "caption": "Fig. 2 Active Length of the SLIM Model",
        "texts": [
            " Hence, the secondary appears to be infinitely long as seen from the primary, and as the secondary rotates in a finite region, hence, within the finite dimensions of any laboratory, it is possible to arrange for continuous motion of the machine, with all constraints satisfied. This type of arrangement is sometimes referred to as \u201cCircular Track LIM\u201d [8]. The specifications of the SLIM are given in Table I. These specifications have been taken from a prototype SLIM present in the Electrical Machines Lab of IIT Kharagpur, India. Fig. 2 zooms the active length of the device. The primary is provided with a three phase full pitch distributed winding (Fig. 3). Each phase has got four series connected coils and each coil has got 18 turns. The nine central slots are double layered and there are three half-filled slots on each side. Each single phase winding can generate four poles, but when altogether three phase supply is given, five poles are created. The primary winding is star connected. The pole pitch is approximately 10cm and the synchronous speed at 50Hz is 10m/sec",
            " On the basis of this information, a set of Non-Linear Simultaneous Equations has been formulated and solved using Optimization Toolbox of MATLAB. As a result, the Equivalent Circuit Parameters have been determined as follows: V. PHYSICAL EXPLANATION OF THE UNBALANCE The following assumptions are essential for the foregoing analysis:  Steady State Analysis has been done; Time transients have been neglected.  Effect of saturation has been neglected.  Anticlockwise direction is the positive direction of motion (Fig. 1 \u2013 Fig. 2).  During positive motion, Slot 1 is the leading edge of the machine, while Slot 15 if the trailing edge (Fig. 3).  The three half-filled slots on either end of the primary does not affect the overall performance of the machine.  The rms value of the flux linking the slot at velocity m/s is denoted by .  The voltage induced in the conductors of the slot at velocity m/s is denoted by . It is known that during motion of a LIM, there is a net decrease in flux at the leading edge of the machine and the flux gradually builds up towards the trailing edge"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003759_gt2016-56951-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003759_gt2016-56951-Figure3-1.png",
        "caption": "FIGURE 3. STEAM TEST RIG",
        "texts": [
            " For this new system, two main requirements have been formulated. The first one is that the new design should include a concept for interchangeable running surfaces. So, different tribopairs between brush seal and rotating counter surface can be investigated without changing the whole rotor system. Second, thermocouples have to be included in the area of brush seals running positions for temperature distribution measurements. A complete schematic overview of the steam test rig, including the new rotor system, is shown in Figure 3. As mentioned above, the general operation principle has not been changed compared to former setups. The central element is the high pressure (HP) chamber, consisting of an annular casing enclosed by cover plates at each end. Designed as a double flow test rig, two opposing brush seal configurations, single or multi-stage, find space within the high pressure chamber. Figure 3 illustrates the schematic view of two multi-stage arrangements. The investigated seal arrangements will be introduced separately. The casing position can be varied relative to the rotor position in order two actively control the seal\u2019s clearance/interference level. The rotor drive concept includes a 80 kW electric motor which is joined to the rotor via a curved teeth coupling. The nominal shaft speed is 10.000 rpm, leading to a maximum surface speed of nearly 160 m/s. The steam is induced in the center of the HP-chamber and is fed back to a power plant steam header after throttling in the individual seal passages"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002031_0142331216631190-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002031_0142331216631190-Figure1-1.png",
        "caption": "Figure 1. Sketch of coning motion.",
        "texts": [
            " One of the main problems of rolling airframe missiles (RAMs) is the cross-coupling between pitch and yaw channels, which reduces the accuracy of the missile motion in the desired direction. In other words, a positive yaw command creates a pitch force in the negative direction and a positive pitch command creates a yaw force in the positive direction. This cross-coupling can be caused by airframe dynamics, actuator response or sensor measurements (Li et al., 2013). This behaviour produces a conical motion along the trajectory (Figure 1). When the radius of the cone is large, the effect of yaw and pitch forces relative to each other increases, and the accuracy of the control system to satisfy the guidance command decreases. Non-linear elements, such as two-position actuators, have undesirable effects on controlling the system. One of the ways to compensate for these effects is injecting a signal with specific features called dither into the input of this element (Zames and Shneydor, 1976). The injection of dither into a non-linear feedback system is widely used in practice to modify non-linearities, to extinguish undesirable limit cycles, to reduce non-linear distortion, to quench jump phenomena, etc"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002655_978-3-319-06590-8_87-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002655_978-3-319-06590-8_87-Figure2-1.png",
        "caption": "Fig. 2 Parameters of pad motion due to perturbation of journal displacement (left) and dynamic pivot properties (right)",
        "texts": [
            " Dimensionful stiffness can be determined by Kik \u00bc jik \u00f02Lg0x\u00de=w3 and corresponding damping by Cik \u00bc cik \u00f02Lg0x\u00de=\u00f0xsw 3\u00de. Thus, both matrices are independent of the frequency ratio. Tilting-pad bearing coefficients The tilting-pad coefficients are determined based on the theory presented by [4] and [14] neglecting the pad mass. Dynamic coefficients of the tilting-pad bearings are derived from the fixed-pad coefficients of each pad. The static shaft displacement Oz,stat is perturbed by De and \u0394\u03b3 according to Fig. 2. Consequently, fluid film forces as well as tilting angles change. The shaft displacement De and \u0394\u03b3 can be transformed to a pad related \u03be\u03b7-coordinate system by: n \u00bc De sin Dc uk\u00f0 \u00de; g \u00bc De cos Dc uk\u00f0 \u00de: \u00f09\u00de Herein \u03c6K is the angular coordinate of the pad pivot. Due to the geometrical proportions the tilting angle \u03b4 can be linearized and expressed by a translational shift of the centre of curvature of the pad: sin d \u00bc x0 Rp \u00fe d d , n0 \u00bc x0 CR \u00bc d Rp \u00fe d CR : \u00f010\u00de Assuming an unrestricted movability of the pad the following equations for the fluid forces can be given by fixed-pad (11) and tilting-pad coefficients (12) considering the modification of tilting angle by \u03be0, and pad deflection \u03b70: DPn \u00bc jng g g0\u00f0 \u00de \u00fe cng g0 g00 \u00fe jnn n n0\u00f0 \u00de \u00fe cnn n0 n00 ; DPg \u00bc jgg g g0\u00f0 \u00de \u00fe cgg g0 g00 \u00fe jgn n n0\u00f0 \u00de \u00fe cgn n0 n00 ; \u00f011\u00de DPn \u00bc 0; DPg \u00bc j gg g g0\u00f0 \u00de \u00fe c gg g0 g00 : \u00f012\u00de If harmonic characteristics of forces and displacements are assumed the equilibrium of fluid forces results in the relationship between tilting-pad and fixed-pad coefficients [14]: j gg \u00bc jgg jnn jgn jng cgn cng \u00fe cnn jgn cng jng cgn j2nn \u00fe c2nn ; c gg \u00bc cgg cnn cgn cng jgn jng \u00fe jnn jgn cng jng cgn j2nn \u00fe c2nn : \u00f013\u00de Consequently, stiffness and damping coefficients of the tilting-pad are each a function of the stiffness and damping coefficients of the fixed-pad. Pursuing this approach to nonsynchronous excitation the dependency of fixed-pad nondimensional damping coefficients on the frequency ratio produces frequency dependent stiffness and damping coefficients of the tilting-pad. In order to consider the influence of dynamic pivot properties the forces effected by pivot stiffness and damping shown in Fig. 2 due to the pad movement n0; n 0 0; g0 and g00 are considered. The quasi-stationary force equilibrium between film and pivot results in the impedance of each tilting-pad. This procedure is comprehensively described by Fuchs [1]. These impedances can be assembled to the eight coefficients of the tilting-pad bearing according to [4] for a certain frequency ratio. In a first step the predictions are validated with comprehensive measurement data presented by Tschoepe and Childs for a flooded lubricated four-pad rocker-pivot, tilting-pad journal bearing [5]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure2-1.png",
        "caption": "Figure 2. (a) Equivalent system with two mutually orthogonal springs without the static force, and (b) equivalent system with two mutually orthogonal springs with the static force.",
        "texts": [
            " Of interest here is the replacement of this concurrent system of springs and dampers by the equivalent, but more simple system that consists of two mutually orthogonal springs and two mutually orthogonal dampers, as shown in Figure 1(b). The task is to find their positions, defined by the angles k and c, respectively, as well as the stiffness coefficients kI and kII, and the damping coefficients cI and cII. The way how to find their characteristics is described in the following subsections. Let us consider now the system of concurrent springs described earlier and shown in Figure 1(a) with a view to finding the equivalent system of two mutually orthogonal springs shown in Figure 2(a). The following theorems and corollaries define this equivalent system. Theorem 1. The system of concurrent pre-stressed linear springs performing small in-plane oscillations can be replaced by the equivalent system of two mutually orthogonal springs on which the original constant static force does not act and which do not have the corresponding static deformations. The position of one of them with respect to the horizontal is given by the angle k, which satisfies the following equation tan 2 k \u00bc Pn i\u00bc1 ki 1 i li sin 2\u2019iPn i\u00bc1 ki 1 i li cos 2\u2019i \u00f01\u00de at NORTH CAROLINA STATE UNIV on May 11, 2015pic",
            " \u00f07\u00de As the system is in the static equilibrium position, the following conditions should be satisfied Fcos \u00f0 \u00de\u00bc Xn i\u00bc1 ki i cos \u2019i \u00f0 \u00de, Fsin \u00f0 \u00de\u00bc Xn i\u00bc1 ki i sin \u2019i \u00f0 \u00de \u00f08a;b\u00de at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from The overall potential energy becomes V \u00bc 1 2 Xn i\u00bc1 ki cos2 \u2019i \u00f0 \u00de \u00fe i li sin2 \u2019i \u00f0 \u00de ( ) x2 \u00fe 1 2 Xn i\u00bc1 ki sin2 \u2019i \u00f0 \u00de \u00fe i li cos2 \u2019i \u00f0 \u00de ( ) y2 \u00fe 1 2 Xn i\u00bc1 ki 1 i li sin 2\u2019i 2 \u00f0 \u00de ( ) xy \u00f09\u00de Note that as the system performs small oscillations around a stable equilibrium position, the potential energy has a quadratic form of the generalized coordinates.12 The angle k which defines the position of one of the spring from the equivalent system (Figure 2(a)) is obtained by equating the coefficient in front of the mixed term xy with zero, which yields Xn i\u00bc1 ki 1 i li sin 2\u2019i 2 k\u00f0 \u00de \u00bc 0 \u00f010\u00de The solution of this equation is equal to that given by equation (1). The expressions in brackets in front of x2 and y2 in equation (9) define the stiffness coefficients kI and kII given by equation (2a,b), which completes the proof of Theorem 1. Corollary 1.1. The system of concurrent linear springs that is not pre-stressed in the stable static equilibrium position and that performs small oscillations can be replaced by the equivalent system of two mutually orthogonal springs, where the position of one of them with respect to the horizontal is given by the angle k, and whose stiffness coefficients are tan 2 k \u00bc Pn i\u00bc1 ki sin 2\u2019iPn i\u00bc1 ki cos 2\u2019i , kI \u00bc Xn i\u00bc1 ki cos 2 \u2019i k\u00f0 \u00de, kII \u00bc Xn i\u00bc1 ki sin 2 \u2019i k\u00f0 \u00de \u00f011a c\u00de Proof",
            " The sum of the stiffness coefficients of the springs from the equivalent systems is given by kI \u00fe kII \u00bc Xn i\u00bc1 ki \u00fe ki i li \u00f012\u00de When the springs from the original system are not pre-stressed in the stable static equilibrium position ( i \u00bc 0), the sum of their stiffness coefficients is equal to the one from the equivalent system kI \u00fe kII \u00bc Xn i\u00bc1 ki \u00f013\u00de i.e. the sum of the stiffness coefficients is the invariant of these systems. Proof. It is sufficient to observe that the sum of the equation (2a,b) gives equation (12) and that for i \u00bc 0, this sum yields equation (13). Let us consider now the system of concurrent springs described earlier and shown in Figure 1(a) with the aim of defining the equivalent system of two mutually orthogonal springs shown in Figure 2(b). Theorem 2. The system of concurrent pre-stressed linear springs performing small in-plane oscillations can be replaced by the equivalent system of two mutually orthogonal linear springs on which the original constant static force also acts and which are, consequently, pre-stressed. The position of one of them with respect to the horizontal is given by the angle k satisfying tan 2 k \u00bc Pn i\u00bc1 ki 1 i li sin 2\u2019iPn i\u00bc1 ki 1 i li cos 2\u2019i \u00f014\u00de where ki is the stiffness of each of the original springs, i is their static deformations, li stands for their corresponding lengths in the static equilibrium position and \u2019i is the angle between each of the spring and the horizontal",
            " The stiffness coefficients k 0 I and k 0 II of these springs are given by k 0 I \u00bc kI F sin k\u00f0 \u00de lII , k 0 II \u00bc kII F cos k\u00f0 \u00de lI \u00f015a; b\u00de where kI \u00bc Xn i\u00bc1 ki cos2 \u2019i k\u00f0 \u00de \u00fe i li sin2 \u2019i k\u00f0 \u00de , kII \u00bc Xn i\u00bc1 ki sin2 \u2019i k\u00f0 \u00de \u00fe i li cos2 \u2019i k\u00f0 \u00de \u00f016a; b\u00de and where lI and lII stand for the lengths of two new springs in the static equilibrium position, which need to be prescribed. at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from Proof. The part of the proof related to the angle k follows from equation (10), as in the case of Theorem 1. However, here the new springs (Figure 2(b)) are pre-stressed and their static deformations can be obtained from the static equilibrium equations Ik 0 I \u00bc F cos k\u00f0 \u00de, IIk 0 II \u00bc F sin k\u00f0 \u00de \u00f017a; b\u00de Equating the potential energies of the springs shown in Figure 2(a) and (b), one can derive kI \u00bc k 0 I \u00fe IIk 0 II lII , kII \u00bc k 0 II \u00fe Ik 0 I lI \u00f018a; b\u00de Combining equations (17a,b) and (18a,b), one can obtain equation (15a,b) as well as the static equilibrium equation (17a,b). Remark. In this approach one needs to define a priori the length of the new springs in the static equilibrium position lI and lII. However, there are other possibilities as well, because the equivalent system contains six unknown quantities k 0 I, k 0 II, lI, lII, I and II that are mutually dependent by means of four equations (17a,b) and (18a,b)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001848_iemdc.2015.7409129-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001848_iemdc.2015.7409129-Figure3-1.png",
        "caption": "Fig. 3: Cross sectional view of the induction machine based EVT.",
        "texts": [
            " The rotor in the center (index 3) is a wound rotor fed by a power electronic converter (PEC) through slip rings. It is connected to the internal combustion engine. The middle rotor is called the interrotor (index 2) and is a short circuited squirrel-cage rotor. It is connected to the final drive of the wheels. Finally the stator (index 1) is a conventional induction machine stator fed by a second PEC. Both PEC\u2019s are connected back-to-back to a common dc-bus. A cross sectional view is shown in Fig. 3. The magnification shows the geometry of the interrotor which consists of alternating iron teeth and aluminum bars. The bars are intersected by an iron flux bridge. This bridge is present only for constructional reasons and is meant to saturate, so that no magnetic short-circuiting takes place. The stator and both rotors are thus magnetically coupled. Some parameters of the studied machine are given in table I. Note that for the interrotor modeling an equivalent wound rotor is considered, meaning that the fundamental m"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003603_1.5118549-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003603_1.5118549-Figure2-1.png",
        "caption": "Figure 2: Experimental setup (front view)",
        "texts": [
            " This paper will show the influence of an adapted intensity distribution, realized by the superposition of two laser beams, on the weld pool geometry resulting in the reduction of spatter formation. . The stainless austenitic steel 1.4541 (X6CrNiTi18 10, Type 321) was used for the experimental trails. The chemical composition is summarized in Table 1. The dimensions of the sheets as well as the position of the weld seam can be found in Figure 1. The specimens were placed in a sample carrier and the pneumatic clamping device (clamping force of 295 N) fixed the plate on one side (see Figure 2). The sample carrier was moved with an industrial robot (KUKA KR 60HA). The experiments were carried out using a Trumpf TruDisk 5000.75 disc laser and a diode laser (LDM 3000, Laserline GmbH). The setup and specifications of the two lasers are summarized in Table 2. Due to angle of incidence for the diode laser, an elliptical spot is depicted on the sample. * measured values with Primes FokusMonitor FM 120 (second moment method) The spot position of the diode laser in relation to the disc laser spot was manipulated by means of translation stages",
            " The angle of incidence was set to 30\u00b0. In order to illuminate the area of interest, the cavilux HF system was used. A narrow band filter for the waverlength of 808 nm was placed in front of the camera. The setup can be seen in Figure 3. The spatter formation was captured with a Nikon DSLR and the position of the camera was perpendicular to welding direction. In order to visualize the droplet escape on the top side as well as on the bottom side, the camera focusses on the free edge of the sample comparable to the front view in Figure 2. The shutter time was set to 1/500 s and the footage was adjusted using the posterize function. In order to quantify the loss of material, the high precision balance (Kern PLJ 2000-3A) was used. The readout amounts 0,001 g and the reproducibility is given with \u00b1 0,002 g. The mass of the sample was determined before and after the welding process in order to calculate the material loss. The experiments were performed three times in order to receive a statistical prediction. The metallographic cross sections were taken in the middle of the sample (see figure 1)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002425_s11771-014-2021-5-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002425_s11771-014-2021-5-Figure4-1.png",
        "caption": "Fig. 4 Discrete surface of cosine gear",
        "texts": [
            " Step 3: Perform automated sequences arrangements of the generating movement parameters and interpolate the discrete movements in a continuous generation motion under the order. The continuous slotting of cosine gear can finally be realized. According to Ref. [16], the equation of the cosine tooth surface can be rewritten as 1 1 1 [ / 2 cos( )]sin [ / 2 cos( )]cos 0 x mz h z y mz h z z            (3) where h denotes the addendum, \u03b8 represents the rotation angle relative to system  (O1, x1, y1) as shown in Fig. 4, m and z denote the modulus and the number of teeth, respectively. According to Fig. 4, assuming that initially the discrete point Mi (i=1, 2, 3, \u2026) on the generated tooth profile at the position (x1i, y1i), the position of the discrete points in the coordinate system (O1, x1, y1) can be expressed as 1 1 1 { /2 cos[ ( )]}sin( ) { /2 cos[ ( )]}cos( ) 0 i i i i i i i x mz h z k k y mz h z k k z                (4) J. Cent. South Univ. (2014) 21: 933\u2212941 936 where \u03b8i is the angle of the tooth profile initially, k is the angle between two adjacent discrete points. The point (x1i, y1i) is one of the discrete points on the digital surface, which is composed of a series of discrete points"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001168_gt2014-26128-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001168_gt2014-26128-Figure6-1.png",
        "caption": "Figure 6: Image showing contours of oil volume fraction in the back chamber for the VOF model. Dark Blue is for =0.01, light blue for =0.1, green for =0.5 and red for =1.",
        "texts": [
            " The jet curves up towards the top of the chamber, splitting on the outer rim and driving the chamber flow from front to back along the outer face and down the back wall towards the oil inlet. Towards the front of the chamber this flow contributes to a small but strong vortex cell at the shroud hole which then leaves through the hole as a jet. 5 Copyright \u00a9 2014 by Rolls-Royce plc Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Figure 6 illustrates the behaviour of the oil phase calculated using the VOF model. As a result of the driving air flow reported above the air flowing down the back wall of the chamber has a significant effect on the oil entering from the bearing as the opposing air flow prevents a film forming and travelling up the back wall. This counter flow contributes to break up of the oil as it enters the back chamber. The VOF model struggles to simulate these conditions and the result is a much dispersed, low volume fraction mist",
            " However it is a very small quantity, around 8x10 -4 kg (around 0.02 litres in the entire back chamber), which is unlikely in a real back chamber. It is therefore unlikely that the oil will ever form a defined film that could be adequately modelled using VOF. It is therefore concluded that VOF is not a suitable multiphase model for this application. The Eulerian model was initialised from the converged VOF simulation. Contours of volume fraction are shown in Figure 8 and as can been seen by comparison with Figure 6, there is a significantly reduced quantity of low volume fraction mist when compare to the VOF model. This mist appears to coalesce and is pushed up to the top wall of the chamber. Here a higher volume fraction film forms. In addition, oil is not immediately broken up upon entering the chamber but instead strips off the back wall in ligaments. These then add to the growing film on the top wall. 6 Copyright \u00a9 2014 by Rolls-Royce plc Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.4-1.png",
        "caption": "FIGURE 8.4",
        "texts": [
            "28) This equation states that the vector sum of the external forces acting on a rigid body is equal to the total mass of the body times the absolute acceleration of the center of mass \u20acxc. The second equation is the rotational equation of motion given as M \u00bc _H (8.29) where the reference point for calculating the applied momentM and the angular momentumH is either fixed in an inertial frame or located at the mass center of the rigid body. If a rigid body rotates about a reference point O, as shown in Figure 8.4, the angular momentum of a small element dV relative to point O is dHO \u00bc x dm _x (8.30) where x is the position vector of the small element relative to the reference point O. dm is the mass of the small element dV. Note that _x is the velocity of dm as viewed by a nonrotating observer translating with O; the velocity is _x \u00bc u x. Therefore, the angular momentum of the rigid body is, with r as the mass density of the body, HO \u00bc Z V rx \u00f0u x\u00dedV (8.31) where u \u00bc uxi \u00fe uyj \u00fe uzk. x \u00bc xi \u00fe yj \u00fe zk. i, j, and k are the unit vectors along the x-, y-, and z-axes, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.8-1.png",
        "caption": "Fig. 4.8 Four-bar mechanism described by relative coordinates",
        "texts": [
            " As far as the \u201crelative coordinates\u201d is concerned, it can be said that this type of coordinates was primarily used in the development of the first general computer programs for mechanisms analysis (Paul and Krajcinovic 1970; Sheth and Uicker 1971). The relative coordinates, also denominated as \u201cjoint coordinates\u201d, define the position and orientation of a body with respect to a preceding body in a multibody system. In general, this type of coordinates is directly associated with the relative degrees-of-freedom allowed by joints that connect bodies. Relative coordinates can be associated with linear or angular displacements, as shown by s and \u03d5 in Fig. 4.7. For the four-bar mechanism of Fig. 4.8, the set of relative coordinates that define its configuration can be stated as q \u00bc /1 /2 /3f gT \u00f04:15\u00de in which the three variables \u03d51, \u03d52 and \u03d53 represent the angle of each body with respect to the x-axis. Since the four-bar mechanism has only one degree-of-freedom, then the three relative coordinates \u03d51, \u03d52 and \u03d53 are not independent, and it is necessary to write a set of two constraint equations. In general, when working with relative coordinates, these equations can be obtained from the closed kinematic chain that defines the configuration of the system. For the case of four-bar mechanism of Fig. 4.8 results that a cos/1 \u00fe b cos/2 \u00fe c cos/3 d \u00bc 0 \u00f04:16\u00de a sin/1 \u00fe b sin/2 \u00fe c sin/3 \u00bc 0 \u00f04:17\u00de Thus, for a given configuration, that is, known for instance \u03d51, the set algebraic equations (4.16) and (4.17) must be solved simultaneously for \u03d52 and \u03d53. This procedure is generally preformed numerically. In summary, the relative coordinates are used to formulate a minimum number of equations of motion of multibody systems. When the system is an open kinematic chain, the number of relative coordinates is equal to the number of degrees-of-freedom, as it is the example of the triple pendulum illustrated in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003014_iraniancee.2016.7585512-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003014_iraniancee.2016.7585512-Figure1-1.png",
        "caption": "Figure 1. Structure of a three phase, 6/4 SRM",
        "texts": [
            " Due to the direct relationship between the winding resistance and temperature, the proposed technique could simply estimate the winding temperature. To confirm the validity of this technique, it is implemented for a three phase, 6/4 SRM using MATLAB software. The estimated winding temperature for different operating conditions including variations in the load torque and speed are presented. The results prove the capability of the proposed EKF based technique. II. STATE SPACE MODEL OF THE SRM In this Section, the state space model of a three phase, 6/4 SRM is presented. Fig. 1 shows the structure of the proposed SRM. As seen, the rotor consists of four poles while the stator includes six poles. The red rectangular parts illustrate the phase windings wound on the stator poles. The state vector for this SRM at the step k is x(k) which could be expressed as: (1) ( ) ( ) ( ) ( ) ( ) ( ) T 1 2 3x k i k i k i k k k\u03b8 \u03c9= In (1), i1, i2 and i3 represent the currents of the three phases. \u03b8 and \u03c9 denote the rotor angular position and the rotational speed, respectively. To estimate the resistance of each phase, one differential equation should be added to the SRM state space equations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000712_s11370-016-0206-5-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000712_s11370-016-0206-5-Figure14-1.png",
        "caption": "Fig. 14 Concept of solving the energy loss",
        "texts": [
            " The tendencies of the contact time with respect to the back swing angle for jokgu ball and basketball are shown in Fig. 13. Though the material properties of the two balls are different, the tendencies of the contact time measured by the FSR sensor and the camcorder were similar to each other. Thus, we can confirm the accuracy of the FSR. In this experiment, the difference of the potential energy between the initial and final positions of the pendulum is taken into account to measure the initial contact velocity \u03c50 M at the distal end of the pendulum. The concept of calculating the energy loss is depicted in Fig. 14. The amount of the energy loss is derived as Eloss = MEgL(1 \u2212 cos \u03b81) \u2212 MEgL(1 \u2212 cos \u03b82) = MEgL(cos \u03b82 \u2212 cos \u03b81) = 1 2 ME (w2 1 \u2212 w2 2) (18) where \u03b81, \u03b82, w1, w2 are the back swing angle, the forward swing angle after the impact (\u03b81 > \u03b82), the angular velocity of the pendulum immediately before and after the impact, respectively. It is assumed that the energy loss is caused by not only the impact but also the friction at the rotating axis. To make the above method reasonable, a smooth bearing is installed at the axis of the pendulum to reduce the friction torque and the retroaction of the pendulum supporter by loading the weight on the hardware"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001782_iros.2015.7354084-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001782_iros.2015.7354084-Figure6-1.png",
        "caption": "Figure 6. Drive unit.",
        "texts": [
            " Human Wall Wall Walking Support Robot (Without Safety Devices) (With Safety Devices) Figure 1. Unexpected high speed robot motion. High Speed Walking Support Robot Human Lock Slope Slope (Without Safety Devices) (With Safety Devices) Figure 3. Case where batteries in the robot have died. Lock Batteries have died. Batteries have died. (Without Safety Devices) (With Safety Devices) Figure 4. Low speed robot motion after the controller has broken down. Large Force Wall Human Wall Lock Contact Force Contact Force-based Safety Device the robot. Fig. 6 shows the drive unit. Each drive unit has a motor with an encoder and the motor torque is transmitted to a wheel via Shaft A, a torque-based safety device, Shaft B, Gear 1-A, Gear 1-B, Shaft C, Gear 1-C, Gear 1-D, and Shaft D. The robot can move by controlling the two motors on the basis of the force sensor signals and the encoder signals. In order to lock Shaft C in clockwise and counterclockwise directions, each drive unit has two velocity-based mechanical safety devices (that is, one velocity-based safety device for locking in the clockwise direction and another velocity-based safety device for locking in the counterclockwise direction)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001082_icicip.2014.7010265-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001082_icicip.2014.7010265-Figure1-1.png",
        "caption": "Fig. 1. Quadrotor axes system.",
        "texts": [
            " Adaptive Techniques in [8], this method gives a good performance with parametric uncertainties and unmodeled dynamics. Sliding Mode approach [9, 10, 11], which is a robust method; however it suffer from the problem of chattering and Backstepping control [3, 12, 13], its main drawback is that is require a lot of computation. II. QUADROTOR MODEL The quadrotor robot is an under-actuated system since it has six Degrees of Freedom (6DoF) while it has only four inputs. The quadrotor has four propellers in cross configuration; two pairs of propellers (front, back) and (right, left) as described in Figure 1, rotate in opposite directions. As we see in Fig.1, we have to consider an inertial frame and a body fixed frame whose origin is in the center of mass of the quadrotor. Thus means that Newton-Euler formalism can be applied. The dynamic model is derived under the following assumptions: 1. Structure is rigid and symmetrical, 2. The center of mass of vehicle and the body fixed frame origin are assumed to coincide, 3. The propellers are rigid in plane, 4. The movement of quadrotor is supposed as a quasi or near quasi flight, 5. For simplicity reasons, we consider that the quadrotor does not exhibit any aerodynamic effect, gyroscopic effect nor friction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure10-1.png",
        "caption": "Figure 10. (a) A frame with a particle of mass m, (b) the force F applied at end point and the coordinates used for defining the internal moments (48a,b) and (c) equivalent two-element system of springs.",
        "texts": [
            "1 is satisfied kI \u00fe kII \u00bc 9984:6 N=m, k 0 I \u00fe k 0 II \u00fe k 0 I I lI \u00fe k 0 II II lII \u00bc 9984:6 N=m \u00f046\u00de The concept presented earlier can also be beneficially applied to the problems containing elastic frames. These problems are related to small oscillations of a particle placed on elastic frames of negligible mass with the aim of obtaining the stiffness coefficients of the equivalent system of two mutually orthogonal springs. Example 3 Let us consider the elastic L-shaped frame of negligible mass shown in Figure 10(a). Its Young\u2019s modulus is E, the second moment of area of the crosssection is I and l stands for the length of each element of the frame. A particle of mass m is fixed at one end as shown in Figure 10(a). To make use of Castigliano\u2019s theorem, the force F is applied at the end point. The values of angle for which the corresponding displacement is extreme (minimal and maximal) will give the directions of the equivalent two-element system of springs. The elastic strain energy due to bending is U \u00bc 1 2EI Z l 0 M1 z1\u00f0 \u00de\u00f0 \u00de 2dz1 \u00fe Z l 0 M2 z2\u00f0 \u00de\u00f0 \u00de 2dz2 \u00f047\u00de where the internal moments at the points defined by the coordinates z1 and z2 in Figure 10(b) are M1 z1\u00f0 \u00de \u00bc z1F cos , M2 z2\u00f0 \u00de \u00bc lF cos \u00fe z2F sin \u00f048a; b\u00de The derivative of the strain energy with respect to the force F is equal to the deflection corresponding to the force, i.e. F \u00bc @U @F \u00bc Fl3 3EI 3 cos2 3 sin cos \u00fe 1 \u00f049\u00de The expression in the brackets has two extreme values, which are I \u00bc 22:5 o, II \u00bc 67:5o \u00f050a; b\u00de The stiffness coefficients of the equivalent two-element system are obtained by substituting the solution (50a,b) into equation (49) kI \u00bc F F I \u00bc F @U @F I\u00f0 \u00de \u00bc 0:649 EI=l3, kII \u00bc F F II \u00bc F @U @F II\u00f0 \u00de \u00bc 7:922 EI=l3 \u00f051a; b\u00de The equivalent two-element system of these springs is shown in Figure 10(c). Example 4 This example is concerned with the arc shown in Figure 11(a), with the particle of mass m attached to at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from its free end. Let us find the angle for which the natural modes of its vibration are directed along the horizontal and vertical direction. First, we introduce a horizontal and a vertical force Q and P (Figure 11(a)) in the place where the particle is located. The elastic strain energy is U \u00bc R 2EI Z 0 M \u00f0 \u00de\u00f0 \u00de 2d \u00f052\u00de where M \u00f0 \u00de \u00bc PR sin \u00f0 \u00de sin \u00bd \u00feQR cos \u00f0 \u00de cos \u00bd \u00f053\u00de According to Castigliano\u2019s theorem @U @Q P 6\u00bc0 Q\u00bc0 \u00bc 0, the angle satisfies the following equation 3 cos2 \u00fe 2 sin cos 2 cos 1 \u00bc 0 \u00f054\u00de Its solution is \u00bc 141:55o \u00bc 2:4705 rad \u00f055\u00de which is shown in Figure 11(b)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002217_icsens.2014.6985116-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002217_icsens.2014.6985116-Figure3-1.png",
        "caption": "Fig. 3. Magnetic Powdery Sensor before adding iron powder (Above) and after adding iron powder (Below).",
        "texts": [
            " 978-1-4799-0162-3/14/$31.00 \u00a92014 IEEE The MPS is a contact-type displacement sensor that consists of iron powder and magnets. Fig.2 is the schematic view of the MPS. Two sets of ABS plates with a magnet fixed on the center are arranged in parallel. The distance between plates varies from 7 mm to 23 mm, and their magnetic fields are in the same direction. We implement the MPS by dispersing iron powder between the plates. Iron powder is attracted to a magnetic field and lines up along the field (Fig.3). When the distance between the two plates is changed, the diameter of the sensing area changes (Fig.4). This diameter change causes a change in the value of the resistance of the sensing area. As a result of this, we can measure a displacement by measuring the resistance value. The key feature of the MPS is that it supports repairing function. The sensing area of the MPS is formed mainly by iron powder particles. Hence, even if the area is broken, the sensing function of the MPS is unaffected as the sensing area is repaired by the supply of iron powder to the area"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000358_9780857095350.6.295-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000358_9780857095350.6.295-Figure1-1.png",
        "caption": "Figure 1: Sketch of the computing modal of crankshaft",
        "texts": [
            " The flywheel and motor were idealized by a set of masses and moments of inertia. The main journal bearings were idealized by a set of linear spring and dash-pots. And in practice modal and forced vibration were analyzed by the finite element method of a 6M51 reciprocating compressor crankshaft system. Crankshaft is considered to be a set of rigidly jointed structures consisting of crank journals crankpins and crank arms. The crank journals and crankpins of each throw were idealized by rods of diameters Dj and Dp and lengths Lj and Lp, as shown in Figure 1. Here Dj and Dp were as equal to their original diameters, while Lj and Lp were equal to the length of the original journal together with the thickness of the crank arm H in accordance with Wlisno and Okmaura (4). The crank arms were idealized by blocks, which dimensions were equal to the original dimensions so as to keep their centers of gravity, masses, moments and stiffness of the same. The flywheel and motor were idealized by a set of masses and moments of inertia about three orthogonal axes attached at their original centers of gravity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001494_amm.732.357-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001494_amm.732.357-Figure1-1.png",
        "caption": "Fig. 1 Deformation of the plate element",
        "texts": [
            "116, University of Auckland, Auckland, New Zealand-01/05/15,00:14:03) The detailed analysis of shell structures is very challenging and still actual. The equidistant distance of points lying between these two surfaces is defined as the shell\u2019s middle surface. The perpendicular length between the surfaces is called the thickness of the shell (denoted h). In praxis, shells have the same characteristics as plates, with one addiction \u2013 curvature. Depending on this fact, shells can be divided into conical, spherical, toroidal, hyperbolic paraboloidal shells and cylindrical [7]. The most widely used theory for laminates are presented in the Fig. 1. The classical laminate theory is based on kinematic assumptions of Kirchhoff-Love theory. The shear theory of first order assumes also shear deformation with comparison to classical laminate theory. This feature enables the application of this theory for the thicker plates [3]. It is assumed that each layer have certain transversal stiffness in the direction of axis 3 \u2261 z. This assumption is used in the relation between stress and strain. After adding the transversal shear stress to the state of the plane stress, we obtain ( ) ( ) ( ) ( ) ( ) ( )                                 =                 \u21d4= 23 13 12 22 11 66 55 44 22 1211 23 13 12 22 11 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001168_gt2014-26128-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001168_gt2014-26128-Figure8-1.png",
        "caption": "Figure 8: Image showing contours of oil volume fraction in the back chamber for the Eulerian model. Dark Blue is for =0.01, light blue for =0.1, green for =0.5 and red for =1.",
        "texts": [
            "org/about-asme/terms-of-use Figure 6 illustrates the behaviour of the oil phase calculated using the VOF model. As a result of the driving air flow reported above the air flowing down the back wall of the chamber has a significant effect on the oil entering from the bearing as the opposing air flow prevents a film forming and travelling up the back wall. This counter flow contributes to break up of the oil as it enters the back chamber. The VOF model struggles to simulate these conditions and the result is a much dispersed, low volume fraction mist. The oil mass in the back chamber is plotted in Figure 8 as a function of time. Here the simulation has clearly reached a balance of oil into and out of the chamber. However it is a very small quantity, around 8x10 -4 kg (around 0.02 litres in the entire back chamber), which is unlikely in a real back chamber. It is therefore unlikely that the oil will ever form a defined film that could be adequately modelled using VOF. It is therefore concluded that VOF is not a suitable multiphase model for this application. The Eulerian model was initialised from the converged VOF simulation. Contours of volume fraction are shown in Figure 8 and as can been seen by comparison with Figure 6, there is a significantly reduced quantity of low volume fraction mist when compare to the VOF model. This mist appears to coalesce and is pushed up to the top wall of the chamber. Here a higher volume fraction film forms. In addition, oil is not immediately broken up upon entering the chamber but instead strips off the back wall in ligaments. These then add to the growing film on the top wall. 6 Copyright \u00a9 2014 by Rolls-Royce plc Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000750_s11668-015-0006-9-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000750_s11668-015-0006-9-Figure2-1.png",
        "caption": "Fig. 2 Schematic diagram of tractor power transmission system and the location of pinion",
        "texts": [
            " Adequate lubrication is necessary to avoid direct metal-to-metal contact between the contacting surfaces and consequent frictional heat. Excessive preload and/or insufficient lubrication has the potential to cause undue rise in temperature due to the increased contact stress between the rib and the roller. In extreme cases, even seizure of bearing can be expected. In the current work, a prematurely failed tail end pinion bearing was investigated. A schematic view of the bearing installation is shown in Fig. 2. A detailed study of the failed bearing was conducted including visual examination, chemical composition, percentage of retained austenite, micro-hardness, and microstructure. The chemical composition of the failed components was carried out using spark emission spectrometry. Bakelite mounted and polished cut sections were used for metallographic observation and micro-hardness measurements. Micro-hardness measurements were carried out using Vickers hardness tester with 0.5 kg load. The microstructure was studied after etching with 3% nital"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003856_j.proeng.2015.05.115-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003856_j.proeng.2015.05.115-Figure2-1.png",
        "caption": "Fig. 2. (a) schematic of the experimental setup; (b) setup inside RMIT Industrial Wind Tunnel.",
        "texts": [
            " A pictorial view and the surface morphology of four balls are shown in Fig. 1. All four balls are FIFA approved. 2.2. Experimental setup RMIT Industrial Wind Tunnel was selected for this study. The tunnel is a closed return circuit wind tunnel with a maximum speed of approximately 150 km/h. The rectangular test section\u2019s dimension is 3 m (wide) 2 m (high) 9 m (long), and is equipped with a turntable to yaw the model. Each ball was mounted on a six component force sensor (type JR-3) as shown in Fig. 2, and purpose made computer software was used to digitize and record all 3 forces (drag, side and lift forces) and 3 moments (yaw, pitch and roll moments) simultaneously. More details about the tunnel and its flow conditions can be found in Alam et al. [8]. A strut support was developed to hold the ball on a force sensor in the wind tunnel, and the schematic of experimental setup with a strut support is shown in Fig. 2. The aerodynamic effect of the strut support was subtracted from the mount with the ball. The distance between the bottom edge of the ball and the tunnel floor was 300 mm, which is well above the tunnel boundary layer and considered to be out of significant ground effect. The aerodynamic drag coefficient (CD) and the Reynolds number (Re) are defined as: AV DCD 2 2 1 (1) VdRe (2) The lift and side forces and their coefficients were not determined and presented in this paper. Only drag data is presented here"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003424_ecce.2016.7854806-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003424_ecce.2016.7854806-Figure5-1.png",
        "caption": "Fig. 5: Comparison of the flux lines between healthy machine and a machine with one magnet fully demagnetized.",
        "texts": [
            " Vd and Vq Variations Under Demagnetization Fault Different reasons may cause demagnetization fault like aging of the magnets, operating the machine under high temperature and field weakening, also in the case of short circuit fault, it might expand and lead to demagnetization of the rotor magnets. Demagnetization faults cause an increase in the torque ripple and reduce the machine efficiency and performance. In the case of demagnetization fault, a nonuniform magnetic flux density will generate around the rotor, which will cause a disturbance to the magnetic flux in the motor and reduce the total magnetic flux density generated from the magnets. Fig. 5 shows a comparison of the magnetic flux line between a healthy machine, and a machine with one magnet demagnetized. The demagnetized region will cause a decreasing in the total permanent flux (\u03bbpm), and based on (3), this will cause a reduction in the total d axis flux. However, The demagnetized region will have the same effect as an air region, forcing the flux to concentrate more in the q axis of the machine. This effect will increase \u03bbq of the machine and decrease \u03bbd. Based on (4) and (5), the magnitude of Vd and Vq will decrease compared with the healthy case"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000567_chicc.2015.7260154-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000567_chicc.2015.7260154-Figure1-1.png",
        "caption": "Fig. 1: The schematic diagram of ball and beam system",
        "texts": [
            " Then, for the sliding motion (8), the transfer function from d to y can be computed as Gyd(s) = C ( sI \u2212 A\u0304\u2212 e\u2212\u03c4sA\u0304\u03c4 )\u22121 (Bd \u2212BKd). According to the final value theorem, to eliminate the disturbance from the output channel, the following condition should be satisfied, C ( A\u0304+ A\u0304\u03c4 )\u22121 (Bd \u2212BKd) = 0. Then, Kd can be solved out as Kd = ( C ( A\u0304+ A\u0304\u03c4 )\u22121 B )\u22121 C ( A\u0304+ A\u0304\u03c4 )\u22121 Bd. (9) 5 Application Examples In this section, we apply the proposed control schemes to stabilize a ball and beam system as shown in Fig. 1. The beam is made to rotate in a vertical plane by applying a torque at the center of rotation, and the ball is placed on the beam, where it is free to roll (with 1 DOF) along the beam. We will assume that the ball remains in contact with the beam and the rolling occurs without slipping, which imposes a constraint on the rotational acceleration of the beam. The parameters used in this system are shown in Table 1. By choosing the beam angle \u03b8 and the ball position \u03b3 as generalized position coordinates for the system, the Lagrangian equations of motion are given by [12] 0 =( Jb R2 +M)\u03b3\u0308 +Mg sin \u03b8 \u2212M\u03b3\u03b8\u03072 \u03c4 =(M\u03b32 + J + Jb)\u03b8\u0308 + 2M\u03b3\u03b3\u0307\u03b8\u0307 +Mg\u03b3 cos \u03b8 where \u03c4 is the torque applied to the beam"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001553_2015-01-1567-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001553_2015-01-1567-Figure3-1.png",
        "caption": "Figure 3. Gyroscopic Torque",
        "texts": [
            " Rack displacement is related to pinion radius using Equation 6 and the total effective mass of the rack includes the rack mass, and the effective translational masses of the tire/wheel assemblies and the electric motor (Equation 7). (6) (7) The rack load is calculated using the following equation: (8) Where, MK/P is the kingpin torque and lknuckle is the kingpin longitudinal offset (knuckle-arm length). The two components for kingpin torque are aligning moment from the tire and the gyroscopic torque due to camber roll. See Figure 3. (9) Tire aligning moment is computed by the product of lateral force and the summation of pneumatic and mechanical trails, which are typically aft of the contact patch center and have negative signs. The gyroscopic torque arises from wheel camber due to vehicle body roll in turn. For a rotating wheel, overturning the wheel (camber) results in a precession (gyroscopic) torque in an axis orthogonal to the axis of wheel rotation and overturning. Equation 10 illustrates this computation: Where, \u03b3roll is camber gain, is roll rate and IWheel-Y is tire/wheel spin inertia"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001962_pesa.2015.7398898-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001962_pesa.2015.7398898-Figure2-1.png",
        "caption": "Fig. 2: Current behaviors during commutation.",
        "texts": [
            " Then, from equations (5) and (6), we can obtain that (3) (4) (5) (6) 3Rt +3L diA -4E = -U (7) A dt m de As the frequency of PWM is very high and the PWM period is much shorter than the electrical time constant LlR, the effect of R can be neglected, the phase currents can be obtained as 4E - U i = -I + m de t A m 3L -U -2E i=l + de m t B m 3L i \" = 2Ude -2Em t C 3L (8) According to equations (2) and (8), the electromagnetic torque during commutation can be calculated as T = 2Em1m + 2Em (U -4 E ) t (9) e 0) 3LO) de m Then the commutation torque ripple can be expressed as U -4E I1T =2 de m E t e 3Lm m According to equation (8), the time that from 1m to 0 is And the time that ie 3LI t == 111 ) Ude +2Em increasing from 0 to 3LI t = m 2 2Ude -2Em (10) iB decreasing (II) 1m IS (12) According to equations (8), (11) and (12), different current behaviors under different speeds are obtained, as shown in Fig.2, and the following conclusions can be also drawn: 1) If Ude < 4Em, then tl < t2 , and the torque keeps decreasing during commutation. 2) If Ude > 4Em, then t) > t2 , and the torque keeps increasing during commutation. 3) If Udc =4Em' then tl = t2, and the torque is constant during commutation. From equation (10), it can obtained that if Ude =4Em during the commutation, the commutation torque ripple will be reduced. From equations (11) and (12), the commutation interval during the commutation can be expressed as III"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003428_icpeices.2016.7853149-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003428_icpeices.2016.7853149-Figure4-1.png",
        "caption": "Fig. 4: Inverted Pendulum System",
        "texts": [
            " The frequency how often crossover is performed depends upon the crossover fraction. Various crossover algorithms are single point, two points, heuristic, and intermediate. Mutation is the occasional random alteration of a value of a string position. The mutation probability is generally kept low. Crossover and mutation are done in order to search the best individual globally. The GA goes on in the above mentioned sequence to give the best solution until the end condition is satisfied. V. INVERTED PENDULUM SYSTEM Figure 4 shows the structure of the Inverted Pendulum system on motor driven cart. The motion of the inverted pendulum is described by two nonlinear equations gives as [6]: .. .. .. mpg)sin (), -mpIZ(), -mp xl cos (), =I(), 1 = mpll{ where, ()I = Pendulum angle from vertical axis Fe = Applied force to the cart x = Displacement of the cart m = I' Pendulum mass Me = Cart mass (9) (10) I = Span between the pendulum center of mass and the pivot point ge = Gravitational constant I p = Inertia of pendulum Using Taylor's series approximation the plant model is linearized about equilibrium point ()I = 0 \u2022 2 Assum ing, sin ()I ::::; ()I , cos ()I ::::; 1, and ()I ()I::::; 0 The transfer function of the Inverted Pendulum system is obtained as follows [6]: -1 u (11) After doing state space modeling of (11), differential equation obtained can be given as [6]: (12) [31 (13) Y\" = XI (14) The values of the plant parameters are given below in Table"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002464_978-981-10-1109-2_5-Figure5.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002464_978-981-10-1109-2_5-Figure5.7-1.png",
        "caption": "Fig. 5.7 The two-step stowage of FRBIO. 1 The deployed position. 2 Partial folding of the side panels inside the unit. 3 Such folded panels further fold inside the unit. 4 Fully stowed unit; VRRFRB\u2212OI = 0.226",
        "texts": [
            " Thismakes such solution particularly suitable for underwater (under-pressured) applications. However, the volume reduction ratio (VRR), that is the relationship of the bounding volumes of the module in stowed (VBs) and deployed (VBd) states, ir relatively high (poor). Although in principle the OI deployment is straightforward and intuitive, in case of super-pressured structures it is not practical. This is due to the linear connections being subjected to tearing. In such cases, the \u201cinside-out\u201d (IO) deployment mechanism, although much more complicated, is more suitable. Figure5.7 shows schematically the two-step stowage of a FRBIO. At first, the side panels partially fold inside. Next such folded panels \u201ccollapse\u201d inside the unit. As Fig. 5.7 indicates, the IO system is substantially more complicated. However, the linear connections are compressed, which is advantageous for super-pressurized structures such as space stations or habitats. Moreover, since the side panels \u201ccollapse\u201d inside the unit, such scheme has better \u201cpacking\u201d capability. Alternatively, the folding of dPZ can be based on collapsing rigid concentric toric rings (CTR). The advantage of this concept is the lack of hinges, as shown in Fig. 5.8. As Fig. 5.8 indicates, CTR is a hingeless system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003907_jrproc.1953.274240-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003907_jrproc.1953.274240-Figure7-1.png",
        "caption": "Fig. 7",
        "texts": [
            " a, - ao (19)1cr2 \\a b2 V--11-----(2A0+ Al) 2 ir D2 E n-ln-B n=l a, an (2A + A1) --(2AO + A3) +- * - r 3a, 4 7r2 A3 -Ao+A1+-+ -+ 4 9 a3 B1-3 -B3\u00b1+ * a, At any turn along the coil, there are two magnetic fields, the self-produced field which is due to the current in the turn itself, and the mutual field which is due to the current in the other turns. The latter may be resolved into two components, axial and radial. (17) A. Self-Produced Field If a unit current is flowing in a certain turn of diameter a, it will be accompanied by a flux linking -2 a 2 Le 1953 541 PROCEEDINGS OF THE 1.R.E. the turn. Assuming a to be increased to a+ba (Fig. 7(a)), 4 will therefore be increased to 4)+&S. The selfinductance I will also be increased to 1+ 31, where 61= i0= ra- B,, 2 and 2 dl B81 - -- = self produced magnetic field due -xa da to unit current. It is considered axial. The mean value of 1 1 dl B.1 - B.Imean = -B81= -- 2 7ra da It is acting over the cross section. The self-produced field due to current i flowing in the turn under consideration will, therefore, be given by i dl Bs8mean ira daa Substitution of (6) gives i 2( d(aAo) d(aAm) (mAr/b) By, 2 db+ cos 2 ra dCa n=m=l dac (21 It should be noted that d(aAm)/da is a function of b/a only, since d(aAAm) b dAm = Am - - da a d(b/a) (20) The value of dl/da is obtained from (3). B. \"Axial Component B,\" of Mutual Field Consider two turns distant y apart and having the same diameter a (Fig. 7(b)). If the diameter of the turn at which the field is required is increased to a+ba, the axial flux linking this turn due to a unit current flowing in the other will be increased from 4v to (Dv+ay The mutual inductance will therefore increase from M to M+6M, where 6a M= bY= ra - Bly, giving 2 dM B, - = - = axial field component ira da due to a unit current passing in the other turn. The mean value of B1i( = B1y) gives the axial field component acting on the cross section. If a current i is flowing in the other turn, 1 2i dM 1 2i dM1 2 ra da 2 -r da C. Radial Component BR of the Mutual Field Let the turn at which the field is required be displaced axially a distance Sy (Fig. 7(c)), the mutual radial flux due to a unit current flowing in the other turn will be decreased from (DR to 4I?R- &R. The mutual inductance will be reduced from M to M-bM, where 1 - M= ra6y. B1R, giving B1R - dM/dy. 7ra The radial field component due to a current i dM i dMl i in other turn = BR = - - --_- _ -ra dy ira dy Substitution of (6) gives i BR = - mAin sin (mry/b). b m=l (22) From (5), (20), (21), and (22), and by following the same lines as in part (3), the axial and radial magnetic field components Bst and Bet, respectively, acting over 542 A pril Mostafa and Gohar: Characteristics of Single-Layer Coils the cross section of a turn distant x will be given by 1 1 2 pb/2 dM Bx= --- i dz 2 D ira J -b/2 da ao 2 b 2 (r d(aAo) 2 r2 a D 2 da + E jn-l 1 d(Asn)Cn n=l1 n da 1 2 b 1 F( d(aAo) a, it, + - _ 2 E sin2 7r2 a D t da n 2 d(aA m) da + E anE,> jm+n-l Cos "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003276_physreve.94.063002-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003276_physreve.94.063002-Figure4-1.png",
        "caption": "FIG. 4. Schematic rendering of a partially inverted cone and its ridge line crease. (a) Pogorelov viewpoint of deformation in which the cone inverts symmetrically under an axial force. The small circles signify the position of the inverting ridge line. (b) Meridional geometry for perfect inversion. The heavy black line is a cone with a sharp apex; the gray line is our cone with a shallow apical cap whose vertical displacement is d . For both, the position of the ridge line is a small circle located a distance b from the edge of the cap and sitting at radius r from the centerline (CL) axis of revolution. The radius of hoopwise curvature at the ridge for a cone is r2, which extends from the centerline to the undeformed meridian immediately below the ridge; \u03b3 is a dummy angle equal to \u03c0/2 \u2212 \u03b1, where \u03b1 is the cone angle. (c) Shown on top is the definition of cap geometry from (b), where R is its radius of curvature; s is an intrinsic coordinate from the top of the cap. The bottom shows a close-up view of the crease geometry around the ridge line. The underlying straight axis is the Pogorelov shape from (a), which intersects the crease ridge line. The local coordinate x is measured from the ridge line on one side only; x = l\u2217 is the characteristic half length of the crease when the transverse displacement w becomes zero again. (d) Half of the crease from (c) with free-body forces and bending moment added to the position of the cut at x = 0. The positive directions of M and Q give dM/dx = Q.",
        "texts": [
            " Although difficult to see, the response is linear at first before softening as the apical cap inverts fully, with the crease forming completely soon after. The crease then rolls down the side of the cone as the applied force decreases exponentially. When the crease is halted by the built-in base, the cone is almost inverted; at this stage, the crease mainly stretches in plane to accommodate d increasing, leading to a rapid rise again in F albeit with a small intermittent dip. At some intermediate displacement, we plot the hoopwise strain \u03b5h in Fig. 3(b) as a function of the intrinsic coordinate s from the apex [Fig. 4(c)], which is the same for every meridian. Moving out of the inverted cap, there are high but localized strains that diminish quickly before increasing again midmeridian at the position of the crease. There is first compressive hoop strain on the inside of the crease facing the inverted apex, followed by tensile strain on the outside. 063002-2 Such compression is surmised as a mechanism for secondary buckling of the crease, but we note a strongly antisymmetrical profile about a local origin that exactly coincides with the ridge line",
            " Such variation is repeated for many conical shapes and thickness, where the radius-to-thickness ratio of shell can range from 15 at the apex to nearly 1300 close to the base. We aim to predict the responsible factors for l\u2217 and so must carefully focus on the kinematic features in our model. However, we also wish to find the force required to pull the crease through the cone. For example, it is not clear why the inverting force should decrease: At the same time, the crease circumference increases, suggesting that more of the cone is becoming strained, viz., more external work and a higher force. We resolve this surmised discrepancy in the following analysis. Figure 4(a) shows a cross section of the equivalent Pogorelov viewpoint of conical inversion. The inverted part can be reflected in the ridge plane to yield the original undeformed cone and all lines are straight. This is the zero strain configuration because there are no changes in length anywhere and zero shell thickness precludes any stored elastic energy in the ridge line. More geometrical detail within the meridional plane is furnished in Fig. 4(b) for a cone with a sharp apex as well as our apical cap. This cone is also perfectly inverted under a polar displacement d with the crease, still of zero width, being located at a distance b along the initial meridian from the end of the cap. Using Figs. 4(b) and 4(c), where the cone angle is denoted by \u03b1 and the cap radius by R, simple geometry affords d = 2R(1 \u2212 sin \u03b1) + 2b cos \u03b1. (1) The current latitudinal radius at the crease position is r , which is equal to R cos \u03b1 + b sin \u03b1. At the same position, the current radius of hoopwise curvature of the conical shell is measured by the distance from the vertical conical axis normal to the meridian, where the usual notation is r2 from [8]. If we select the meridian in the upright part, then r2 is drawn as per Fig. 4(b) and r/r2 = cos \u03b1. An explicit expression for r2 in terms of d can now be found by eliminating b between them: r2 = R + tan \u03b1 cos \u03b1 [ d 2 \u2212 R(1 \u2212 sin \u03b1) ] . (2) This returns r2 = d tan \u03b1/2 cos \u03b1 for a perfect cone when R is zero. Strictly speaking, the radius of curvature of the actual ridge line is not defined because of the discontinuity in gradient there and the expression for r2 is only true immediately on either side of the ridge (if we had selected the inverted meridian instead, then r2 changes direction and lies above r). However, this presents no problem because at the ridge line itself, we need to interrogate the local crease shape in order to ascribe curvature properties in detail. Figure 4(c) shows a schematic axisymmetrical cross section of a crease, which is drawn relative to the Pogorelov outline. Displacements away from this outline constitute straining and curving, where the normal component w is assumed to be dominant. We also assume that w is symmetrical about the ridge line because the crease is far enough away from the base or apex. The formulation can now be contracted slightly by defining a positive meridional coordinate x moving away 063002-4 from the ridge towards the base; x does not take negative values, rather, any functions of it are merely reflected about x = 0 either symmetrically or antisymmetrically provided we are careful about boundary conditions at x = 0",
            " Beyond the ridge line, r2(x) increases, but if w decreases sharply with x then there is little change in r2 within the crease itself: In other words, the radius of hoopwise curvature of the crease is everywhere 1/r2, with r2 expressed by the current value of Eq. (2). As noted, meridional strains are not negligible, but w does not depend on them directly. The curvature, however, in this direction \u03ba is set by the shallow gradient expression \u03ba = \u2212d2w/dx2, where absolute differentials assert the onedimensional nature of problem. We idealize further by excising half of the crease in Fig. 4(c) for all of x and redrawing it in Fig. 4(d). At the cut x = 0 and thus everywhere on the ridge line, there must be a bending moment M and locally aligned forces in shear Q and in tension T ; these are all axisymmetrical quantities defined per unit length of shell and are drawn in their positive directions. Even though we do not show the far field forces at the base in this view, it is possible to write certain equilibrium relationships between them and M , Q, and T . However, specified in this way, our problem turns out to be statically indeterminate and we must invoke arguments of geometrical compatibility for a complete solution",
            " The corresponding auxiliary equation sets \u03be 4 + 4\u03b24 = 0, returning roots of \u03be = (\u00b11 \u00b1 \u221a\u22121)\u03b2 and a general solution of four terms as A1e \u03b2x sin \u03b2x + A2e \u2212\u03b2x sin \u03b2x . . . . Those terms involving exponential growth can be discounted for w rapidly decaying with x, resulting in two terms whose amplitudes are found by setting M = D\u03ba and Q = dM/dx at x = 0: w = e\u2212\u03b2x 2\u03b22D [ M(sin \u03b2x \u2212 cos \u03b2x) + Q \u03b2 cos \u03b2x ] for x 0. (4) Ensuring that w = 0 at x = 0 obviates M = Q/\u03b2, giving a gradient of M/2\u03b2D at the same position, which must equal \u03c0/2 \u2212 \u03b1 from Fig. 4(d). The final expression for w turns out to be rather compact: w = (\u03c0/2 \u2212 \u03b1) \u03b2 e\u2212\u03b2x sin \u03b2x for x 0. (5) Because of the proportionality between w and hoop strain \u03b5h and given the latter\u2019s variation in Fig. 3(b) on either side of the ridge line, we could have inferred the expression above by inspection. The formal analysis is conclusive and Eq. (5) immediately offers l\u2217 when we set \u03b5h, and hence w, equal to zero at x = l\u2217, giving sin \u03b2l\u2217 = 0 and \u03b2l\u2217 = \u03c0 . Recalling our previous definitions of \u03b2 and D, we can therefore write l\u2217 = \u03c0 [ 4Dr2 2 Et ]1/4 \u21d2 l\u2217 = \u03c0r 1/2 2 t1/2 [3(1 \u2212 \u03bd2)]1/4 (l\u2217 \u2248 2",
            "4, giving a crease length 2l\u2217, just slightly less than r2. However, the ratio quickly attenuates to around 0.1\u20130.15, which implies that r2 changes little over the crease span, thereby confirming our original proposal. Figure 6 deals with different cone angles of the same thickness and, again, the differences between theory and computational data are minimal. Even though a steeper cone implies a slower rate of change of r2 along a meridian, which bolsters the cylindrical shell analogy, the initial gradient in Fig. 4(d) at x = 0 also increases, which undermines the shallow shell assumption. For a tube where \u03b1 = 0\u25e6, this gradient is infinite yet Eq. (6) differs by less than 10% compared to finite 063002-5 element data, which is remarkable. In this case, r2 is a constant and equal to the radius of the original apical cap. Note that in general, as the cone angle decreases, so does l\u2217. We now proceed to finding the inverting force F using an energy formulation for the entire cone based on the strain energy stored in the crease and not in the deformed apical cap"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000758_0885328215585682-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000758_0885328215585682-Figure1-1.png",
        "caption": "Figure 1. Structure of phospholipid bilayer.",
        "texts": [
            "5,6 Among the many great properties of graphene are its remarkably high stiffness with Young\u2019s modulus of approximately 1000GPa, its outstanding heat conductivity of 3000W (mK) 1, its high-speed electron mobility of 200,000 cm2 (V s) 1 at room temperature and high specific surface area of 2600m2 g 1.7\u201310 Therefore, graphene would represent a new revolutionary asset in material science, nanoelectronics and matter physics.11 Moreover, compared to any metal used on biosensors such as biomimetic membrane-coated types, graphene allows current to pass through it in a much more precise and effective manner and at higher speed. In many plant and animal cells, there is a two-layer covering called phospholipid bilayer surrounding the cell. Figure 1 depicts phospholipids, the molecules which form phospholipid bilayer.12 These molecules can organize themselves by forming a pair of layers which form a covering that only some specific types of substances are merely able to infiltrate. Such a structure provides a barrier preventing unwanted materials from entering the cell.13 1Institute of High Voltage and High Current, Faculty of Electrical Engineering, Universiti Teknologi Malaysia, Johor Bahru, Malaysia 2Department of Communications Engineering, University of Sistan and Baluchestan, Zahedan, Iran 3Department of Electrical Engineering, Neyriz Branch, Islamic Azad University, Neyriz, Iran 4Department of Electrical Engineering, Islamic Azad University, Yasooj, Iran Corresponding author: Zolkafle Buntat, Institute of High Voltage and High Current, Faculty of Electrical Engineering, Universiti Teknologi Malaysia, Johor Bahru 81310, Malaysia"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002380_transjsme.15-00563-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002380_transjsme.15-00563-Figure3-1.png",
        "caption": "Fig. 3 Hinge design detail",
        "texts": [],
        "surrounding_texts": [
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n2. \u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u958b\u767a\n2\u00b71 \u554f\u984c\u63d0\u8d77\n\u672c\u7814\u7a76\u3067\u306f\uff0cUAV\u3068\u3057\u3066\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u7528\u3044\u308b\uff0e\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u8907\u6570\u306e\u30d7\u30ed\u30da\u30e9\u306b\u3088\u308a\u63a8\u529b\u3092\u5f97\u308b\u3053\u3068\u3067\u6d6e \u4e0a\u3057\uff0c\u305d\u308c\u305e\u308c\u306e\u30d7\u30ed\u30da\u30e9\u306e\u63a8\u529b\u6bd4\u3092\u5236\u5fa1\u3059\u308b\u3053\u3068\u3067\uff0c\u6a5f\u4f53\u306e\u59ff\u52e2\u5236\u5fa1\u3092\u884c\u3046\uff0e\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u56fa\u5b9a\u7ffc\u6a5f\u3068\u6bd4 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\u3053\u3068\u304c\u78ba\u8a8d\u3067\u304d\u308b\u304c\uff0c\u5e02\u8ca9\u54c1\u306f\uff0c\u4e00\u822c\u306b\uff0c\u98db\u884c\u524d\u306b\u4eba\u529b\u3067\u6a5f\u4f53\u306e\u5c55\u958b\u4f5c\u696d\u3092\u884c\u3046\u5fc5\u8981\u304c\u3042\u308b\uff0e\u305d\u3053\u3067\uff0c\u672c\u7814\u7a76\u3067",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u306f\u3053\u308c\u306b\u52a0\u3048\uff0c\u8155\u90e8\u306e\u6298\u308a\u305f\u305f\u307f\u65b9\u5411\u3068\u30d2\u30f3\u30b8\u69cb\u9020\u306b\u7740\u76ee\u3059\u308b\u3053\u3068\u3067\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306b\u3088\u308b\u81ea\u5df1\u5c55\u958b\u304c\u53ef\u80fd\u306a\u69cb \u9020\u3092\u63d0\u6848\u3059\u308b\uff0e\u63d0\u6848\u3059\u308b\u69cb\u9020\u3067\u306f\uff0c\u5c55\u958b\u30fb\u56fa\u5b9a\u306e\u305f\u3081\u306e\u98db\u884c\u306b\u76f4\u63a5\u95a2\u4fc2\u7121\u3044\u30a2\u30af\u30c1\u30e5\u30a8\u30fc\u30bf\u304c\u4e0d\u8981\u3068\u306a\u308a\uff0c\u6a5f\u4f53\u91cd \u91cf\u3092\u6291\u3048\u3089\u308c\u308b\uff0e\n\u63d0\u6848\u306b\u57fa\u304d\u8a2d\u8a08\u3057\u305f\u30d2\u30f3\u30b8\u90e8\u306e\u8a73\u7d30\u3092\u56f3 3\u306b\u793a\u3059\uff0e\u3053\u306e\u30d2\u30f3\u30b8\u306f\u30b9\u30e0\u30fc\u30ba\u306b\u53d7\u52d5\u52d5\u4f5c\u3059\u308b\u305f\u3081\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u304c \u52d5\u4f5c\u3057\u3066\u3044\u306a\u3044\u5834\u5408\uff0c\u91cd\u529b\u306b\u3088\u308a\uff0c\u8155\u90e8\u306f\u4e0b\u306b\u6298\u308c\u66f2\u304c\u308b\uff08\u56f3 3-\u5de6\uff09\uff0e\u4e00\u65b9\uff0c\u30d7\u30ed\u30da\u30e9\u306e\u63a8\u529b\u65b9\u5411\u3068\uff0c\u8155\u90e8\u306e\u5c55\u958b \u65b9\u5411\u3092\u4e00\u81f4\u3055\u305b\u308b\u3053\u3068\u306b\u3088\u308a\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u3092\u56de\u8ee2\u3055\u305b\u3066\u63a8\u529b\u3092\u5f97\u308b\u3053\u3068\u3067\uff0c\u8155\u90e8\u304c\u81ea\u7136\u306b\u5c55\u958b\u3059\u308b\uff08\u56f3 3-\u4e2d\uff09\uff0e \u8155\u90e8\u306f\uff0c\u6a5f\u4f53\u30d9\u30fc\u30b9\u90e8\u3068\u540c\u4e00\u5e73\u9762\u4e0a\u3068\u306a\u308b\u5730\u70b9\u3067\uff0c\u30b9\u30c8\u30c3\u30d1\u306b\u3088\u308a\u5c55\u958b\u304c\u5236\u9650\u3055\u308c\uff0c\u4e00\u822c\u7684\u306a\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u5f62\u72b6 \u3068\u306a\u308b\uff08\u56f3 3-\u53f3\uff09\uff0e\u8155\u90e8\u306e\u7a81\u8d77\u3068\u30b9\u30c8\u30c3\u30d1\u304c\u5d4c\u307e\u308a\u5408\u3046\u3053\u3068\u3067\uff0c\u30d2\u30f3\u30b8\u90e8\u306e\u904a\u3073\u306b\u3088\u308b\u8155\u90e8\u306e\u306d\u3058\u308c\u30fb\u632f\u52d5\u3092\u9632\u3050\uff0e\n\u306a\u304a\uff0c\u672c\u8ad6\u6587\u3067\u306f\u6298\u308a\u305f\u305f\u307f\u6a5f\u69cb\u306e\u6709\u7528\u6027\u3092\u691c\u8a3c\u3059\u308b\u3053\u3068\u3092\u76ee\u7684\u3068\u3059\u308b\u305f\u3081\uff0c\u6a5f\u4f53\u306b\u306f\u5b9f\u969b\u306e\u63a2\u67fb\u306b\u5229\u7528\u3059\u308b\n\u30bb\u30f3\u30b5\u3084\uff0c\u5916\u58c1\u3068\u306e\u63a5\u89e6\u9632\u6b62\u7528\u306e\u30d7\u30ed\u30da\u30e9\u30ac\u30fc\u30c9\u7b49\u306f\u642d\u8f09\u3057\u306a\u3044\uff0e\n2\u00b72\u00b71 \u8155\u90e8\u306e\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u306e\u691c\u8a3c\n\u8155\u90e8\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u3092\u691c\u8a3c\u3059\u308b\uff0e\u56f3 4\u306b\u53ef\u52d5\u8155\u90e8\u306b\u50cd\u304f\u529b\u3092\u793a\u3059\uff0e\u3053\u306e\u56f3\u306b\u304a\u3044\u3066\uff0c\u30d2\u30f3\u30b8\u306e\u56de\u8ee2\u4e2d\u5fc3\u3067\u3042\u308b\n\u70b9 Q\u56de\u308a\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u306e\u5f0f\u306f\u6b21\u5f0f\u306e\u3068\u304a\u308a\u3068\u306a\u308b\uff0e\u305f\u3060\u3057\uff0cl f \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u30d7\u30ed\u30da\u30e9\u4e2d\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0c l f g \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u53ef\u52d5\u8155\u90e8\u91cd\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0cm\u3092\u53ef\u52d5\u8155\u90e8\u306e\u8cea\u91cf\uff0cFm \u3092\u30d7\u30ed\u30da\u30e9\u63a8\u529b\uff0c\u03b8 f \u3092\u53ef\u52d5\u8155\u90e8\u306e \u6298\u308a\u305f\u305f\u307f\u89d2\u5ea6\uff0cg\u3092\u91cd\u529b\u52a0\u901f\u5ea6\u3068\u3059\u308b\uff0e\u306a\u304a\uff0c\u7c21\u5358\u306e\u305f\u3081\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306f\u8155\u90e8\u4e0a\u3067\u529b\u3092\u767a\u63ee\u3059\u308b\u3082\u306e\u3068\u3059\u308b\uff0e\n0 =\u2212l f gmgcos\u03b8 f + l f Fm (1)\n\u3053\u306e\u5f0f\u3092 Fm \u306b\u3064\u3044\u3066\u307e\u3068\u3081\u308b\u3068\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm = l f g\nl f mgcos\u03b8 f (2)",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3057\u3066\u3044\u308b\u5834\u5408\uff0c\u03b8 f = 0\u3067\u3042\u308b\u305f\u3081\uff0c\u53ef\u52d5\u8155\u90e8\u3092\u5b8c\u5168\u5c55\u958b\u72b6\u614b\u306b\u4fdd\u3064\u305f\u3081\u306e\u6761\u4ef6\u306f\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm \u2265 l f g\nl f mg (3)\n\u4e00\u65b9\u3067\uff0c\u6a5f\u4f53\u304c\u30db\u30d0\u30ea\u30f3\u30b0\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u6a5f\u4f53\u306e\u5168\u8cea\u91cf\u3092\u652f\u3048\u308b\u3060\u3051\u306e\u63a8\u529b\u304c\u5fc5\u8981\u3068\u306a\u308b\uff0e\u30db\u30d0\u30ea\u30f3\u30b0\u6642\uff0c4\u3064 \u306e\u30d7\u30ed\u30da\u30e9\u304c\u3059\u3079\u3066\u540c\u3058\u63a8\u529b\u3092\u767a\u63ee\u3057\u3066\u3044\u308b\u3068\u4eee\u5b9a\u3059\u308b\u3068\uff0c\u6a5f\u4f53\u5168\u4f53\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u304b\u3089\u30d7\u30ed\u30da\u30e9\u3042\u305f\u308a\u306e\u63a8\u529b Fm \u3092\u6c42\u3081\u308b\u3068\u6b21\u5f0f\u3067\u8868\u73fe\u3067\u304d\u308b\uff0e\u305f\u3060\u3057\uff0cmb \u3092\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u8155\u90e8\u4ee5\u5916\u306e\u6a5f\u4f53\u8cea\u91cf (\u30d9\u30fc\u30b9\u90e8\u8cea\u91cf)\u3068\u3059\u308b\uff0e\n4Fm = (mb +4m)g\nFm = ( 1 4 mb +m ) g (4)\n\u3053\u306e\u5f0f\u3068\uff0c\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3059\u308b\u305f\u3081\u306e\u6761\u4ef6\u5f0f (3)\u304b\u3089\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306b\u53ef\u52d5\u8155\u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u306a\u3044\u305f\u3081\u306e\u6761 \u4ef6\u304c\u6b21\u5f0f\u3068\u306a\u308b\uff0e( 1 4 mb +m ) g \u2265 l f g l f mg\nmb \u2265 4m ( l f g\nl f \u22121\n) (5)\n\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3067\u306f\uff0c\u53ef\u52d5\u8155\u90e8\u306e\u5148\u7aef\u306b\u6700\u91cd\u91cf\u7269\u306e\u30e2\u30fc\u30bf\u304c\u3064\u3044\u3066\u304a\u308a\uff0c\u305d\u306e\u70b9\u3067\u63a8\u529b\u304c\u767a\u63ee\u3055\u308c\u308b\uff0e\u305d\u306e \u305f\u3081\uff0cl f g < l f \u3068\u306a\u308a\uff0c\u5f0f (5)\u306e\u53f3\u8fba\u306f\u5e38\u306b\u8ca0\u3068\u306a\u308b\u3053\u3068\u304b\u3089\uff0c\u5e38\u306b\u6210\u308a\u7acb\u3064\uff0e\u3053\u308c\u3088\u308a\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u98db\u884c\u6642\u306b\u8155 \u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u3053\u3068\u306f\u7121\u3044\u3053\u3068\u304c\u8a3c\u660e\u3055\u308c\u305f\uff0e\n2\u00b72\u00b72 \u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u6642\u306e\u52d5\u4f5c\n\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306f\u3059\u3079\u3066\u306e\u30d7\u30ed\u30da\u30e9\u304c\u307b\u307c\u304a\u306a\u3058\u56de\u8ee2\u6570\u30fb\u63a8\u529b\u3092\u767a\u751f\u3057\u3066\u3044\u308b\u304c\uff0c\u79fb\u52d5\u306e\u305f\u3081 \u306b\u306f\u5404\u30d7\u30ed\u30da\u30e9\u3067\u63a8\u529b\u5dee\u3092\u4f5c\u308b\u3053\u3068\u3067\u6a5f\u4f53\u3092\u50be\u3051\uff0c\u6c34\u5e73\u65b9\u5411\u306e\u63a8\u529b\u3092\u5f97\u308b\uff0e\u307e\u305f\uff0c\u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u306f\uff0c\u5404\u30d7\u30ed\u30da\u30e9\u306b \u3088\u308a\u767a\u751f\u3059\u308b\u53cd\u30c8\u30eb\u30af\u3092\u5229\u7528\u3057\u3066\u5b9f\u73fe\u3059\u308b\uff0e\u305d\u306e\u305f\u3081\uff0c\u79fb\u52d5\u6642\u3084\u30e8\u30fc\u56de\u8ee2\u6642\u306b\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u3088\u308a\u3082\u63a8\u529b\u304c\u5c0f \u3055\u304f\u306a\u308b\u30d7\u30ed\u30da\u30e9\u304c\u5b58\u5728\u3059\u308b\u3053\u3068\u306b\u306a\u308a\uff0c\u524d\u8ff0\u306e\u8155\u90e8\u3092\u4fdd\u3064\u6761\u4ef6\u3092\u6e80\u305f\u3059\u3053\u3068\u304c\u3067\u304d\u306a\u3044\u5834\u5408\u304c\u5b58\u5728\u3059\u308b\uff0e\u7279\u306b"
        ]
    },
    {
        "image_filename": "designv11_64_0002325_1475090215624721-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002325_1475090215624721-Figure1-1.png",
        "caption": "Figure 1. Vessel model and its reference frames.",
        "texts": [
            " While each type of control system has its benefits for certain kinds of applications, this study is concerned with the design of novel vessel DP systems using mixed HN and msynthesis through a reduced-order controller. Particularly, the simple weighting functions are used to eliminate first-order wave energy from the control signals, while allowing only second-order lower frequency energy to remain. Finally, the presented robust control schemes are relatively simple, and provide computationally efficient algorithms under the large parametric variations as well as LF and HF disturbances with measurement noises. As illustrated in Figure 1, the vessel dynamics can be described using two frames of reference that are important for DP: an Earth-fixed (inertial) frame, {E}:= [OEXEYEZE] and a vessel-fixed (body) frame, {B}:= [OXYZ]. Assume that the vessel has an XZplane of symmetry, and Z-axis is pointing out of paper. The frame {E} has origin at OE, and OG is the center of gravity (CG) of the vessel with position (xG, yG, zG). The reference frame {E} will be used for global positions and all commanded positions. The frame {B} is free to translate and rotate in the frame {E}; xG denotes at The University of Melbourne Libraries on June 5, 2016pim"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002723_978-3-319-45447-4_62-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002723_978-3-319-45447-4_62-Figure2-1.png",
        "caption": "Fig. 2. Types of specimens and their geometry",
        "texts": [
            " Also, exposure to UV radiation can cause stress concentrators on the surface of the polymeric material (which can lead to a premature failure), and can make the material more sensitive to the action of other factors (i.e. humidity) [10]. The paper evaluates the changes in physical and mechanical properties of composites reinforced with hemp fabric as a result of thermal and ultraviolet degradation. The analyzed composite material contain hemp mat and polyurethane resin (RAIGITHANE 8274/RAIGIDUR CREM), with 50 % percentage of reinforcing natural fibers. To evaluate the rheological behavior of composite reinforced with hemp fibers, from the plate were cut 5 samples coded with the letter C (Fig. 2) whose geometric and physical characteristics are shown in Table 1. This material is a commercial product, currently used in car body structure. Initial, the characteristic curve (\u03c3-\u03b5) and elastic modulus were determined at tensile test. So, the average values of elastic modulus are: E1 = 2773 MPa (in longitudinal direction) and E2 = 3770 MPa (in the transverse direction). To study the rheological behavior of the composite materials reinforced with mat hemp and polyurethane resin in the first stage was examined the roughness of surfaces specimens using atomic forcemicroscope (AFM -Atomic ForceMicroscope) NTEGRAProbe Nanolaboratory, in the way semi-contact, using a cantilever type NSG 10 with constant force from the Institute for Research and Development of Transilvania University of Brasov"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003908_1.1698360-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003908_1.1698360-Figure1-1.png",
        "caption": "FIG. 1. Principal arrangement of the bearing.",
        "texts": [
            " At this very limit of the axial compression a slight sidewise bending occurs perpendicular to that axis of the cross section about which the moment of inertia is the least. Within certain narrow limits neither the amount nor the sign of that lateral deflection is defined. This means that within those limits any deflection of the column's center part can occur with practically no additional bending force. This effect, suitably trans ferred to the design of a bearing, can be used to yield a rotation within a small range of angle without.applying any force for turning. Figure 1 shows the principal arrangement of such a bearing. On a circle A, with the radius Rl attached tion are concerned. Two different types of beam bearings are treated, and their equations for preca1culation are derived. Furthermore, the influences of temperature, external forces, and the angle of rotation on the torque of the bearing are investigated. Finally, a combination of two different beam bearings is shown to offer a possible compensation for these unfavorable effects. to the movable part of the bearing, there are three metal rods rigidly secured in the radial directions. The other ends of those metal beams are fixed on the circle B belonging to the stationary part of the bearing. The distance of these ends from the center of motion 0 of the bearing is R2\u2022 The three rods are somewhat displaced in the axial direction of the bearing so that they will not disturb each other. In stressing one of them by the spring S (Fig. 1) all three are under equal tensile stress. A suitable choice of the tensile force, p., can produce a phenomenon similar to the above men tioned of a column under axial compression. This means that in the case of the bearing the restoring moment between the two circles A and B can be come zero or even negative. Physically it can be explained as follows: By turning of circle A the fixed ends of the metal beams move on a circle with 297 [This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to ] IP: 141.209.100.60 On: Sun, 21 Dec 2014 14:43:11 the radius Rl (see dash line in Fig. 1, marking the position of one of those metal beams). It is evident that because of the tensile force of the beams a negative restoring moment acts on the circle A and increases with the angle of rotation. On the other hand, the beams give a positive restoring moment because of their natural rigidity. Therefore it seems evident that for a suitable choice of all proportions, especially of P\" the total restoring moment of the bearing can be made zero or even negative. The following discussion gives a short quantita tive representation of the relations wherefrom all consequent requirements for the design of such bearings can be derived",
            " This change in length of the elastic curve between the points I and II results in a decrease of P., the consequence of which, accord ing to Fig. 3, is a positive restoring torque. Ob viouslysuch a bearing has only a negligible re storing torque in the immediate surroundings of VOLUME 20, APRIL, 1949 rp=O, which, however, increases with cp to positive values as P. varies. These disadvantages can be nearly compensated for if a second bearing under compression stress is used. The principal arrangement of such a bearing exactly equals the one shown in Fig. 1. The only difference is that the spring S supplies a compression force. The differential equation of the elastic curve of such a single beam is (d2y/dx2) + (Px / IE)\u00b7y = - (Py / IE) \u00b7x+ CMt/ IE). (12) With regard to the boundary conditions, the solution can be written as follows: y = (Py/Px) [(A/cosa) . sin \u00abx/\"A) +a) -x->.\u00b7tana], (13) where A represents again A=(EI/P,,)O .\u2022. The angle a is defined by the relation Rt/;\\.+sin(L/;\\.)+Rd;\\.\u00b7cos(L/;\\.) tana 1 -cos(L/}.) + Rd}.\u00b7 sin (L/}') (14) (15) The total torque of point I wi th respect to the center of motion 0 can be expressed by COSa' (R l - >"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000387_978-1-84996-080-9_7-Figure7.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000387_978-1-84996-080-9_7-Figure7.12-1.png",
        "caption": "Fig. 7.12 Model (a) and real vehicle (b) with ROV operation buoyancy configuration \u2013 neutrally or marginally positive buoyancy",
        "texts": [
            " However, the PC104 can be accessed and remotely administered from the Control PC through the network using the remote administrator. 2 Ship, ROV, ocean surface, seabed and so forth The Thrusted Pontoon/ROV is a multi-mode of operation vehicle. It can be operated on the surface (Figure 7.11) as a survey platform either towed (by an extra towing cable) or thrusted by 4 horizontal thrusters to allow surge, sway and yaw which is useful in confined spaces or near hazards where a boat and tow cannot operate. It can also be operated as an ROV (Figure 7.12). In these various modes of operation, it is used in conjunction with an umbilical and associated winch; the umbilical carrying vehicle power, control and data from sensors and instruments. In the surface-tow or surface thrusted modes of operation, overall vehicle buoyancy is maintained strongly positive by 8 buoyancy modules mounted on the pontoon upper frame. While in surface-tow mode, a \u201cQuick Release\u201d arrangement allows the two top most buoyancy modules to be detached from the vehicle, reducing overall vehicle buoyancy to neutral or slightly positive"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure3.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure3.7-1.png",
        "caption": "Fig. 3.7 \u201cZero-order\u201d generator",
        "texts": [
            "3); that is, r0N \u00bc rN\u00f0Ri r1N\u00de rN \u00f0Ri r1N\u00de : \u00f03:9\u00de The resistances Ri, r0N have the following characteristic values: RV i \u00bc 0; rV0N \u00bc rNr1N rN \u00fe r1N ; \u00f03:10\u00de which defines the generalized equivalent generator as an ideal voltage source; RI i \u00bc 1; rI0N \u00bc rN ; \u00f03:11\u00de which defines the generalized equivalent generator as an ideal current source; R0 i \u00bc RG L \u00bc r1N \u00fe rN ; r00N \u00bc 1; \u00f03:12\u00de which corresponds to the beam G0 and defines the \u201czero-order\u201d source when the current and voltage of the load are always equal to zero for all the load values. The generalized equivalent generator that displays the \u201czero-order\u201d generator is presented in Fig. 3.7. VL = 0 because the internal resistance voltage Vi 0 = \u2212 VL G. R i R I i =\u221e R V i =0 r V 0N V G L r 0N R L I G L 0 V L G I L R 0 i r 0 0N =\u221e r I 0N Fig. 3.6 Family of the load straight lines with the characteristic values Ri and r0N 3.2 Circuit with a Series Variable Resistance 59 So, a variable element and internal resistance can have these three specified characteristic values. These values are defined at a qualitative level. This brings up the problem of determination in the relative or normalized form of the value r0N regarding of these characteristic values"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure2.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure2.7-1.png",
        "caption": "Fig. 2.7 (a) Pure shear state of plane stress. (b) Unit of stress under the shear deformation",
        "texts": [
            " For a more general consideration, a local cubic volume inside a material body may have stresses in all x-, y- and z- directions, so the Cartesian strains can be expressed as \u03b5x \u00bc \u03c3x \u03bd\u03c3y \u03bd\u03c3z =E \u03b5y \u00bc \u03c3y \u03bd\u03c3z \u03bd\u03c3x =E \u03b5z \u00bc \u03c3z \u03bd\u03c3x \u03bd\u03c3y =E \u00f02:17\u00de Apparently, for the shear stresses and strains in the three Cartesian directions \u03b3x \u00bc \u03c4x=G, \u03b3y \u00bc \u03c4y=G, and \u03b3z \u00bc \u03c4z=G \u00f02:18\u00de Both G and E are quantities for measuring the stiffness of materials under different modes of external forces. In fact, these moduli are dependent on each other and also the involvement of \u03bd. For homogeneous isotropic linear elastic materials, we consider a cubic material element to be subjected only to shear as shown in Fig. 2.7a. The external applied shear stresses are \u03c4xy (i.e., the shear stress in the y-direction acting on the surface with an x-normal axis) and \u03c4yx (acting along the x-direction on the surface with a y-normal axis) where |\u03c4xy| \u00bc |\u03c4yx|. Recall Eq. 2.17: \u03b5x \u00bc 1 E \u03c3x \u03bd \u03c3y \u00fe \u03c3z \u00f02:19\u00de 45 in the xy-plane. If we choose the new Cartesian coordinates by rotating the axes 45 counterclockwise about the z-axis such that the new x-axis (x0) aligns with maximum stress and the new y-axis (y0) aligns with the minimum stress: \u03b50 \u00bc \u03b5max \u00bc 1 E \u03c4xy \u03bd 0 \u03c4xy \u00bc \u03c4xy E 1\u00fe v\u00f0 \u00de \u00f02:20\u00de 42 2 Basic Material Properties This strain, which deforms the element along the new x0 axis, can also be related to the shear strain \u03b3xy. Suppose the side length is dx in the plan element OABC, and the left side OA is fixed, as shown in Fig. 2.7b. Due to shear deformation, the right side BC with a length of dx dislocates with the distance BB0 \u00bc \u03b3xydx, where \u03b3xy \u00bc \u03c4xy/G. If we now consider the new coordinate frame with OB as the principal axis with the strain \u03b5max, the length of OB is ffiffiffi 2 p dx, and its elongation is BD with a length as 1= ffiffiffi 2 p times that of BB0, i.e., \u03b3xydx= ffiffiffi 2 p . Therefore, the maximum strain is \u03b5max \u00bc \u03b3xy/2 \u00bc \u03c4xy/(2G). \u03b5OB \u00bc \u03b5max \u00bc \u03b3xy 2 \u00bc \u03c4xy 2G \u00f02:21\u00de Substituting Eq. 2.17 into Eq. 2.16 and rearranging the terms, we can obtain the relationship between Poisson\u2019s ratio (v), modulus of rigidity (G), and modulus of elasticity (E) as G \u00bc E 2 1\u00fe \u03bd\u00f0 \u00de \u00f02:22\u00de that thermal strain \u03b5T explains this change due to any temperature change for a 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000817_tmag.2013.2287030-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000817_tmag.2013.2287030-Figure2-1.png",
        "caption": "Fig. 2. Branching points and BH curves.",
        "texts": [
            " (1) 2) The branching points are given by the extrema of H . When H begins to decrease (or increase) after increasing (decreasing), the local maximum (minimum) of H gives the new branching point, which becomes the initial point; the former initial point becomes the terminal point. 3) When H passes through Hf , the branching points Pi and Pf are deleted and the previous two branching points return as the initial and terminal points. This property is similar to the wiping-out or deletion property governing the Preisach model [2], [3]. Fig. 2 shows branching points Pk at (Hk, Bk) (k = 1, \u2026) and P\u00b1S at \u00b1(HS, BS), where PS and P\u2212S are positive and negative saturation points. The probability function for the descending BH curve from P1 is defined by the branching points P1 (initial point) and P\u2212S (terminal point), which is denoted by p\u2212(H : P1, P\u2212S). When H begins to increase at P2 before reaching P\u2212S , P2 becomes a new branching point (initial point) and the ascending BH curve (second order reversal curve) from P2 toward P1 (terminal point) is determined by the probability function p+(H : P2, P1)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure7-1.png",
        "caption": "Fig 7 eddy current density distribution on no-load",
        "texts": [],
        "surrounding_texts": [
            "2 1 0\n0 1\n1\n( , ) cos( ) cos( )\n1 {cos[( 1) ( ) ] 2\ncos[( 1) ( ) ]\ne r r e\nr e r\ne r\nb t F p t t\nF p t\np t\n  \n  \n\n       \n     \n    \n   \n    \n    \n\n (12)\nThe slip relative to rotor is 1 2 1 1 es p        ( ) , so the flux\ndensity relative to rotor is\n2 2_ 0 1\n2\n2 1\n2\n1 ( , ) {cos[( 1) + ( ) ]\n2 1\ncos[( 1) + ( ) } 1\ne rotor r e r\ne r\ns b t F p t\ns\ns p t\ns\n  \n\n      \n    \n    \n    \n\n(13) 3) Fundamental component of armature magnetic field\n3 1 1 1 0\n1 0 1 1\n1 1\n( , ) cos( ) cos( )\n1 {cos[( 1) ( ) ] 2 cos[( 1) ( ) ]}\ne s s e\ns e s\ne s\nb t F p t t\nF p t\np t\n      \n    \n   \n   \n    \n    \n(14)\nThe harmonic order is 1\n1 p  .The slip relative to rotor\nis 1 3 1 1 es p       ( ) , so the flux density relative to rotor is\n3 3 _ 1 0 1 1\n3\n3 1 1\n3\n1 ( , ) {cos[( 1) ( ) ]\n2 1\ncos[( 1) ( ) ] 1\ne rotor s e s\ne s\ns b t F p t\ns\ns p t\ns\n     \n   \n     \n     \n(15)\n4) Harmonic component of armature magnetic field\n4 1 0\n0 1\n1\n( , ) cos( ) cos( )\n1 {cos[( 1) ( ) ] 2\ncos[( 1) ( ) ]\ne s s e\ns e s\ne s\nb t F p t t\nF p t\np t\n  \n  \n\n       \n     \n    \n   \n    \n    \n\n (16)\nThe slip relative to rotor is 1 4 1 1 es p        ( ) , so the flux\ndensity relative to rotor is\n4 4_ 0 1\n4\n4 1\n4\n1 ( , ) {cos[( 1) + ( ) ]\n2 1\ncos[( 1) + ( ) ]} 1\ne rotor s e s\ne s\ns b t F p t\ns\ns p t\ns\n  \n\n      \n    \n    \n    \n\n(17) The combination of equations (11), (13), (15) and (17) leads to the flux density relative to rotor for a machine with an eccentric rotor.\n_ 1_ 2_\n3_ 4_\ne rotor e rotor e rotor\ne rotor e rotor\nb t b t b t\nb t b t\n  \n \n \n \n( , ) ( , ) ( , )\n( , ) ( , ) (18)\nCalculate the additional eddy current losses in permanent magnet which is caused by eccentric rotor. The analytical modeling showed in Fig3 is based on the following assumptions: a) Neglect end effect; b) Neglect eddy current reaction; c) The magnetic field in air gap only has radial component.\nBased on polar coordinate 2-D model, the magnetic vector potential in permanent magnet can be given as\n_e rotorA r Bd r b t d     ( , ) (19)\nThe eddy current in the magnet can be calculated by A\nJ t\n   \n (20)\nWhere is the conductivity of magnet. The combination of equations (19) and (20) leads to the eddy\ncurrent density in magnets.\n1 1_ 2 2_\n3 3_ 4 4_\ne rotor e rotor\ne rotor e rotor\nJ r s b t s b t\ns b t s b t\n  \n \n  \n \n[ ( , ) ( , )\n( , ) ( , )] (21)\nThe magnet loss caused by eccentric rotor can be calculated by\n2 2 2 3 3 0\n2 2 2 2 2 2 2 2 1 1 2 3 1 4\n1\n1\n24\ne eddy\nb a\nr r s s\np L J J dS\nL R R\ns F s F s F s F\n\n   \n\n   \n \n \n   \n\n \n*\n_\n( )\n( )\n(22)\nWhere L is the active length of magnet.  is the width of\nmagnet. aR , bR are the inner radius and external radius of magnet, respectively. For the different kinds of air gap eccentricities, the components in harmonic magnetic field are different. For the\nstatic eccentricity 0e  , 1 3\n1\n1 s s p\n  \n , 2\n1\n1 s p    ,\n4 1 1\np s\np    . For the dynamic eccentricity e  ,\n1 2 3 0s s s   , 4 1\n1\np s\np    .\nBased on the analysis above, it comes to a conclusion that both rotor magnetic field and armature magnetic field can induce eddy current loss in magnet under static eccentricity. For the dynamic eccentricity, only armature harmonic magnetic field will induce",
            "eddy current loss in magnet. In additional, the eddy current loss is proportion to the square of eccentricity  , whether static eccentricity or dynamic eccentricity.\nIII. NUMERICAL CALCULATION\nThe eddy current loss in carbon fiber sleeve is very small due to the low conductivity. The influence of carbon fiber sleeve on magnetic field is negligible. So the effect of eccentricity on magnet will be calculated both on static eccentricity and dynamic eccentricity.\nA. The influence of static eccentricity on rotor loss\nThe flux density waveform on magnet surface due to static eccentricity and the corresponding FFT analysis are shown in Fig 5. The flux density on the eccentric side increases and on the opposite flux density decreases. From the point of the stator, the air gap along the circumference doesn\u2019t vary with rotor angle, but the length of air gap for each magnet varies with rotor angle, which leads to the alternating magnetic field on magnet surface. The alternating magnetic field will induce additional eddy current loss on rotor. Fig 6 gives out the harmonic order on the surface of magnet due to static eccentricity. Compared with noneccentric rotor, the main harmonic order is 1/2and 3/2, which agrees with the equation (10).\nTable 1 gives the rotor loss when the static eccentricity is 0.6. Both magnet loss and rotor yoke loss are more than non-\neccentric rotor. The total loss increases nearly 30 times, but the loss on each magnet increases almost the same.\nThe eddy current density distribution is compared to a noneccentric rotor in Fig 7and Fig 8. It has been seen that with 0.6eccentricity, the eddy current density has significant changes.\nAccording to the equation (22), it is obvious that the eddy current loss on magnet is proportion to the square of eccentricity .\nFig 9 shows that the loss increases with the increase of eccentricity. From the above analysis, it is seen that static eccentricity has a visible effect on magnet loss. Hence we should",
            "try to avoid large static eccentricity and make better efficiency and lower magnet loss.\nB. The influence of dynamic eccentricity on rotor loss\nIn the case of dynamic eccentricity, the position of minimum air gap length rotates with the rotor position making the eddy current loss on magnet different with static eccentricity. The rotating axis is off-center from rotor axis. The effect of dynamic eccentricity varies with different eccentric angle. Eccentric angle is the included angle between rotating axis and rotor axis as Fig 10 show. So the effect of eccentric angle on magnet will be calculated on dynamic eccentricity.\nSince for most of the additional harmonic magnetic field, rotor slip is zero under dynamic eccentricity, so whether no-load or load will not cause large additional loss in the rotor. Healthy and faulty machines with different eccentric angle under different load conditions were simulated. The result is shown that dynamic eccentricity only increases a little rotor loss compared with healthy machine in Fig 11.The effect caused by eccentric angle is also very small.\nBut as Fig 12 shows the different eccentric angle will affect the distribution of rotor loss. In a healthy machine, the eddy current loss in each magnet is equal. When the dynamic eccentric angle =0\u00b0, PM1 has the minimum air gap length, which leads to the maximum harmonic magnetic field making eddy current loss increase. PM3 has the maximum air gap length, so the eddy current loss on PM3 is the smallest. PM2 and PM4 are symmetric, so they have the same eddy current loss. Fig 13 gives\nFig 14 shows that the rotor loss at different eccentricity ratio when eccentricity angle is zero. It is seen that dynamic eccentricity has a small effect on rotor loss. But the loss will focus on PM1 which will make it produce larger temperature rise. Therefore, we should try to avoid dynamic eccentricity."
        ]
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.55-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.55-1.png",
        "caption": "Fig. 2.55 Three-element truss structure with different external loading: a force boundary conditions; b displacement boundary conditions",
        "texts": [
            " The three truss elements have the same cross-sectional area A and Young\u2019s modulus E . The length of each element can be taken from the dimensions given in the figure. The structure is loaded by (a) a horizontal force F at node 2, (b) a prescribed horizontal displacement u at node 2. Determine for both cases \u2022 the global system of equations, \u2022 the reduced system of equations, \u2022 all nodal displacements, \u2022 all reaction forces, \u2022 the force in each rod. 2.36 Plane truss structure arranged in a triangle Given is the two-dimensional truss structure as shown in Fig. 2.55. The three truss elements have the same cross-sectional area A and Young\u2019s modulus E . The length of each element can be taken from the dimensions given in the figure. The structure is loaded by 86 2 Rods and Trusses (a) single forces FX and FY at node 4, (b) prescribed displacements u X and uY at node 4. Determine for both cases \u2022 the global system of equations, \u2022 the reduced system of equations, \u2022 all nodal displacements, \u2022 all reaction forces, \u2022 the force in each rod. 2.5 Supplementary Problems 87 of each element can be taken from the dimensions given in the figure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.21-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.21-1.png",
        "caption": "FIGURE 6.21",
        "texts": [
            " In order to develop a full vehicle model based on roll stiffness it is necessary to determine the roll stiffness and damping of the front and rear suspension elements separately. The estimation of roll damping is obtained by assuming an equivalent linear damping and using the positions of the dampers relative to the roll centres to calculate the required coefficients. If a detailed vehicle model is available, the Determination of front end roll stiffness. procedure used to find the roll stiffness for the front suspension elements involves the development of a model as shown in Figure 6.21. This model includes the vehicle body, this being constrained to rotate about an axis aligned through the front and rear roll centres. The roll centre positions can be found using the methods described in Chapter 4. The vehicle body is attached to the ground part by a cylindrical joint located at the front roll centre and aligned with the rear roll centre. The rear roll centre is attached to the ground by a spherical joint in order to prevent the vehicle sliding along the roll axis. A motion input is applied at the cylindrical joint to rotate the body through a given angle",
            " The road wheel parts are not included nor are the tyre properties. The tyre compliance is represented separately by a tyre model and should not be included in the determination of roll stiffness. The wheel centres on either side are constrained to remain in a horizontal plane using inplane joint primitives. Although the damper force elements can be retained in the suspension models they have no contribution to this calculation as the roll stiffness is determined using static analysis. The steering system, although not shown in Figure 6.21, may also be included in the model. If the steering system is present, a zero-motion constraint is needed to lock it in the straight-ahead position during the roll simulation. Front end roll simulation. For the rear end of the vehicle the approach is essentially the same as for the front end, with in this case a cylindrical joint located at the rear roll centre and a spherical joint located at the front roll centre. For both the front and rear models the vehicle body can be rotated through an appropriate angle either side of the vertical"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure3-1.png",
        "caption": "Fig. 3. Three Examples of Articulated Floating Shoes",
        "texts": [
            " The first group might aptly be described as \u201cpivoted\u201d; a word which would then have to be avoided in relation to the floating types. \u201cNon-floating\u201d is not an attractive title, although adequate. \u201cAnchored\u201d is a possible alternative, but as ships float when at anchor it would be a misnomer if applied to shoes which do not float. \u201cRigid\u201d has been used, but is unsuitable since the shoe may deliberately be given flexibility. The following nomenclature is therefore proposed :- (1) Pivoted shop (Fig. 1). (2) Floating shoes : (a) Sliding (Fig. 2). (b) Articulated (Fig. 3). Other items are defined as follows :- Leading shoes: shoes in which the movement of the Trailing shoes: shoes in which the movement of the * An alphabetical list of references is given in Appendix 11, p. 51. drum over the lining is towards the pivot or abutment. drum over the lining is towards the applied load. Toe and heel: the ends of the lining in relation to the \u201cankle\u201d at the pivot or abutment. Shoe tip : the part of the shoe at which the operating load is applied; that is, the end remote from the pivot or abutment",
            " In some designs a very strong spring is used in order to provide a large frictional force, but it can be mounted in such a position as to keep its moment about the abutment down to a value which wi l l not make the brakes heavy to operate. Sliding shoes and their abutments can probably be produced and serviced more economically than any other type, and this alone could account for the trend in this direction, and perhaps for over-emphasis of their functional advantages, material as they are. at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from Articulated Shoes. The Huck type (Fig. 3a) of articulated floating shoes has very similar characteristics to those of a sliding shoe with the abutment face at right angles to the centreline of the link; the slotted shoe (Fig. 3b) belongs in the same category, but the Lockheed tank type (Fig. 3c) yields to the same analysis of forces as a pivoted shoe except that concentricity is guaranteed. One of the advantages of the articulated type over the sliding shoe is that by applying a considerable amount of friction at the joint between the shoe and the link, and none at the anchor pin, the shoe can be located, when the brakes are released, exactly as if it were a pivoted shoe. The trend, however, is towards sliding shoes, and while articulated shoes have been used extensively and successfully, they can for the purposes of this classification be regarded as hybrids, the main issue being a comparison between pivoted and sliding types",
            " The conclusions which can be drawn from the table are that the sliding shoe tends to have a lower output than the pivoted type and to be more stable, but that it is possible to obtain a very similar performance with either type of shoe, by suitable selection of 2 and h. I = 1.15 and K = 0.75. -drae shoetip load\u2018 F = Pivotedshoes . . . . . . 11- atrailing Sliding shoes with parallel abutments . . . 3leading 4trailing 5 leading 6trailing Sliding shoes or Huck type with 21 deg. abutments at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from INTERNAL EXPANDING SHOE The factors of floating shoes of the Huck type (Fig. 3a) and those with sliding abutments inclined to one another (Fig. 2b) cannot be expressed as simple formulae but they are easily obtained by graphical methods. They are plotted in Fig. 6 for a typical shoe, of which the shoe-tip and abutment forces act at 21 deg. to each other, together with the factors of pivoted shoes, and sliding shoes with parailel abutments, with I = 1.15 and k = 0.75 in each case. The relative inclination of the abutments does not affect the value of p at which the leading shoe will sprag, which will occur when cosece = l / k for all types of floating shoe. This is because the shoe-tip load becomes virtually non-existent at the spragging point. (Racipmcslsofshoefpctors..) I = 1-15 and k = 0.75. Pivoted shoes . . . . . . 1 leading 2 trailing Sliding shoes with parallel abutments . . . 31eading 4 trailing 5 leading 6 trailing Sliding shoes or Huck type with 21 deg. abutments In the Lockheed tank type (Fig. 3c), spragging occurs when cot 6 = Z/k where k is taken, as it must be for the factor analysis, as the distance OB from the brake-drum centre to the pivot of the carrier. At the spragging point, however, the shoe-up load disappears and the carrier becomes completely analogous to the link of the Huck shoe, and if k is now measured as OC, the perpendicular distance to the line through the two articulations of the shoe, then cosec 6\u2019 = Z/k as for the other floating types. It will be noted that the factor of the trailing shoes is considerably less subject to the p-value than that of the leading shoes, the equivalent value of Fo"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003027_978-3-319-46532-6_7-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003027_978-3-319-46532-6_7-Figure1-1.png",
        "caption": "Fig. 1 Schematic picture of the F-IVT architectures",
        "texts": [
            " In practical applications, it would be made of two devices in series: a fixed ratio drive and a continuously variable transmission. In particular, the transmission will be of infinitely variable transmission type (IVT) in order to permit also the change of the rotation wise of the output shaft. The fixed ratio drive is a harmonic drive (HD) or a ball screw (BS), depending on the attachment to the joint/leg. In the former case, the F-IVT is of rotating type, in the latter case it is of linear type (Fig. 1). For one given walking regime and in ideal conditions, (negligible mechanical power losses and flywheel with infinitely large inertia), the variable power request of the knee would be filtered by the flywheel, and the motor should only feed the system with the average power needed to walk, applying also an almost constant torque to the flywheel, thus working on a fixed operating point at optimal electric efficiency. Thus, since average power and torque are much smaller than their peak values the motor can be greatly undersized"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure10-1.png",
        "caption": "Figure 10. Shaft-lock mechanism. (d)",
        "texts": [
            " As the velocity of Gear A (i.e. Shaft C) increases, the damping torque increases. Claw D rotates by the torque difference between the damping torque and the spring torque, and locks Plate A, if the velocity of Shaft C exceeds the detection velocity level. The detection velocity level is adjustable by changing the attachment position of Linear Spring D as shown in Fig. 9. We can change the attachment position of Linear Spring D by using an adjustment mechanism of the detection velocity level as shown in [13]. Fig. 10 shows the mechanism to mechanically lock Shaft C. After Plate A is locked, each Claw B slides along each Guide Hole A of Plate A by the rotation of Plate B (Fig. 10(b)) and then one of three Claws B is hooked to the inner teeth and rotates Plate C (Fig. 10(c)). By the rotation of Plate C, Pin C switches off and each Claw C moves along each Guide Hole C (Fig. 10(d)). After that, four Claws C simultaneously mesh with Ratchet Wheel C and thus Shaft C is locked. The lock of Shaft C is released by rotating Shaft C in a direction opposite to the direction locked by the safety device. The torque-based safety device consists of a commercial torque limiter (Fig.11, TGB20-HC, Tsubakimoto Chain Co., see [14] for more details) and a switch (Switch Z). As shown in Fig. 12, Switch Z is installed at the position of being pressed by Plate Z of the torque limiter when Plate Z slides along Hub of the torque limiter",
            " 16 and Table II, we consider that the velocity-based safety device switched off the motor and locked Shaft C after the velocity of Shaft C approximately became the detection velocity level. The differences between the detection velocity levels and the experimental values are attributed to the damping torque errors of Rotary Damper, the attachment position errors of Linear Spring D, etc. Furthermore, in Fig. 16, the difference between the time when Claw D locked Plate A and the time when Shaft C was locked is attributed to the presence of the process of \u201cPlate A\u2019s stop (see Fig. 9 and Fig. 10(a))\u201d, \u201cClaw B\u2019s rotation (Fig. 10(b))\u201d , \u201cPlate C\u2019s rotation by Claw B (Fig. 10(c))\u201d, \u201c\u2018switch-off by Pin C\u2019 and \u2018Shaft C\u2019s lock by Claws C\u2019 (Fig. 10(d))\u201d which is required in the shaft-lock mechanism of the velocity-based safety device. The above experimental results indicate that the torque-based safety device can switch off the robot\u2019s motor, if the torque-based safety device detects the unexpected high torque even when the unexpected high velocity does not occur in the drive-shaft. Also, they indicate that the velocity-based safety device can switch off the robot\u2019s motor and lock the drive-shaft, if the device detects the unexpected high velocity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002373_21681015.2016.1174162-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002373_21681015.2016.1174162-Figure3-1.png",
        "caption": "Figure 3.\u00a0schematic diagram of the neutral line of the flexspline.",
        "texts": [
            " Here, \u03d5 is the angle between the rotational axis Y1 and the fixed axis Y2. The moving coordinate {O1, X1, Y1} is fixed with flexspline TP. \u03b3 is the polar angle of tooth root of the flexspline. \u03bc is the angle between the symmetric line of the flexspline TP to line OO1. \u03c6 is the angle between the input terminal (wave generator) and output terminal (back face of the flexspline), \u03c61 is the angle between long axis of the wave generator and the line OO1, \u03c62 is the angle of the long axis of wave generator to axis Y2. The geometric relation is shown in Figure 3, (5) \u23a7 \u23aa\u23a8\u23aa\u23a9 xE = \u2212 \ufffd R + rw \ufffd sin \ufffd 2rw 0 R+2rn \ufffd + rw sin \ufffd 2 0(R+rw) R+2rw \ufffd yE = \ufffd R + rw \ufffd cos \ufffd 2rw 0 R+2rw \ufffd \u2212 rw cos \ufffd 2 0(R+rw) R+2rw \ufffd (6) \u23a7 \u23aa\u23a8\u23aa\u23a9 xF = \ufffd R \u2212 rn \ufffd sin \ufffd 2rn 0 R\u22122rn \ufffd \u2212 rn sin \ufffd 2 0(R\u2212rn) R\u22122rn \ufffd yF = \ufffd R \u2212 rn \ufffd cos \ufffd 2rn 0 R\u22122rn \ufffd + rn cos \ufffd 2 0(R\u2212rn) R\u22122rn \ufffd (7) \u23a7\u23aa\u23aa\u23aa\u23a8\u23aa\u23aa\u23aa\u23a9 x 2 = X 1 cos + Y 1 sin + sin y 2 = \u2212X 1 sin + Y 1 cos + cos x 2 t \u22c5 y 2 \u2212 x 2 \u22c5 y 2 t = 0 = + = 1 \u2212 2 D ow nl oa de d by [ U ni ve rs ite L av al ] at 0 7: 14 1 2 M ay 2 01 6 conjugate characters. With the parameters in Table 1, substitute Equation (3) into (7), the theoretical TP of circular spline is solved and shown in Figure 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003124_urai.2016.7734077-FigureI-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003124_urai.2016.7734077-FigureI-1.png",
        "caption": "Fig. I. In-pipe robot integreted with all the modules",
        "texts": [],
        "surrounding_texts": [
            "Nowadays pipelines are necessary condition for transporting gas, oils and fluids in most industrial facilities. When pipelines aging, they were suffered defects. To prevent the expected disaster, it is necessary to regular inspection on these facilities. However, because pipelines buried under the ground, it is difficult to be accessed. For this reason the in-pipe robot have been developed for inspection and maintenances. Most of the in-pipe robots are difficult to pass through elbow, branch pipe and miter except the straight pipe. Thus, multi modular type robots [I] were equipped with joint between robot components that improve mobility in pipe elements like Fig.!. But, when robots pass through small curvature pipe and sharp edges inside pipe, the path of in-pipe robot should be complicated. If the path becomes more complex, the size of the robot should be smaller and it take considerably long time in order to pass through elbow. It is difficult to simply pass through the pipe element even with the joint module. In this paper, control strategy was proposed using the joint module. Joint module between robot components will be able to pass through the various elements of the pipe. In particular, it was described the joint module control strategy to simply pass through elbow. 2. CONTROL STRATEGY 2.1. Under development 1n-pipe robot Driving module of in-pipe robot has been developing and has been applied differential gear mechanism, it is not nec essary to control the pass through elbow [2]. Joint module of in-pipe robot has back-drivability, so it doesn't have to control [3]. However, a front sensor module and a rear sensor module are different from the other module, there is no equipment for supporting. Therefore, the joint module behind sensor module, it is necessary to separate control. 2.2. Scenario As shown in Fig.2 the steps how to track reference trajectory 2-pitch joint module can pass through the elbow. Consequently angle variations of 2-pitch joint module are presented ac cording to the following calculation formula. There are four steps for determining the reference trajectory. Step 1 and 2 determines the angle relative to the direction of between driving module and starting line of elbow. Step 3 and 4 is related to a camera module and ending line of elbow. The movements of the joint module is calculated by consid ering x. Xl is distance between driving module side pitch axis and stating line of elbow, X2 is camera module side pitch axis and ending line of elbow. Where Ct, {-3 are angles of 2-pitch , a 978-1-5090-0821-6/16/$31.00 \u00a92016 IEEE 442 is the length between pitch axis and center of camera module, b is the length between pitch axis and center of driving module, c is the distance between 2-pitch axis and D is the diameter of pipe. In the step 1, the angular variation is represented in Eq.(1) and (2). cos -1 l.5D J(l.5D)2 + xi (1) Following the above formula is continued until when Xl is zero. In the step 2, the angular variation is represented in Eq.(3) and (4). o xi + c2 - a2 a = 270 -cos-1:----7;::::==; =;=:::;=;:;==:::::;c 2cJ(l.5D)2 + xi COS-1-r\ufffd\ufffd X \ufffd1\ufffd=\ufffd J(l.5D)2 + xi (3) Following the above formula is continued until the center of the rear module goes up to the starting line of the elbow. The angle of the moment enables to be infinite rotation in circular elbow. A sensor module was moved to the ending line while maintaining the angle of the moment. Next, step 3 and 4 can be considered as reverse order of the step 1 and 2. And Xl is changed to X2. The angle variation is calculated in accordance with the following equation. In the step 3 and 4 the angular variation is represented in Eq. (5)-(8)."
        ]
    },
    {
        "image_filename": "designv11_64_0003763_insi.2015.57.5.283-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003763_insi.2015.57.5.283-Figure3-1.png",
        "caption": "Figure 3. Spur gear tooth modelling: (a) XY plane[3]; (b) YZ crosssection",
        "texts": [
            " (7) Considering the gear tooth as a non-uniform cantilever beam, the strain energy can be represented as a decomposition of bending, shear and axial compressive energy as a result of the equivalent force F. The energy due to bending deformation can be calculated by: Ub = Mz 2I y + 2MzMyI yz + My 2Iz 2E Iy Iz \u2212 I yz 2( )0 d \u222b dx ................. (8) where E is the elastic modulus. My,Mz are the internal moments and Iy,Iz,Iyz are the second moments of area of the cross-section dx, as represented in Figure 3. The energy due to shear deformation is calculated by: Us = Fb 2 2GIz 2 1 t y( ) yt y( )dy y ytop\u222b\u23a1\u23a3\u23a2 \u23a4 \u23a6\u23a5\u2212 ybottom ytop\u222b 2 dy \u23a7 \u23a8 \u23aa \u23a9\u23aa \u23ab \u23ac \u23aa \u23ad\u23aa dx 0 d \u222b ....... (9) where G is the shear modulus and t(y) is the sectional width at the distance y from the neutral axis, as shown in Figure 3. The energy due to axial compressive deformation is calculated by: Ua = Fa 2 2EA x( ) dx0 d \u222b .............................. (10) where A(x) is the area of the cross-section dx. Fb and Fa are calculated by: Fb = F cos \u03b1( ),Fa = F sin \u03b1( ) ..................... (11) d and \u03b1 are shown in Figure 3. Based on Equations (7) to (11), we can obtain the expressions Kb,Ks,Ka, which are, respectively, the local bending, shear and axial compressive stiffness. The stiffness of the Hertzian contact between the teeth pair can be approximated by[5]: Kh = \u03c0EW 4(1\u2212\u03bd 2 ) ...................................... (12) The expression of the stiffness is constant along the pressure line, which is independent to the position of contact and the depth of interpenetration[5]. The stiffness value is dependent on the width of contact between two teeth W and the material properties, Poisson\u2019s ratio and elastic modulus, \u03bd and E, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003125_systol.2016.7739793-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003125_systol.2016.7739793-Figure2-1.png",
        "caption": "Fig. 2: Inverted pendulum on a carriage.",
        "texts": [
            " However, accurate trajectory tracking is only possible for the uncertain system with an active variable-structure part. Analogously, this holds for the second-order case which is omitted due to page constraints. IV. INTERVAL-BASED VARIABLE-STRUCTURE CONTROL OF AN INVERTED PENDULUM As a second application, the interval-based variable structure controller is applied to the stabilization of an inverted pendulum, with an output error constraint formulated in terms of a two-sided barrier function. The inverted pendulum depicted in Fig. 2 consists of a carriage of mass M on which a pendulum of length a is mounted. In good accuracy, the pendulum can be assumed to consist of a massless rod while its mass m is concentrated in the tip of the pendulum. The horizontal carriage displacement is denoted by z (t), the angle of the pendulum by \u03b1 (t), and the accelerating force acting onto the carriage by FC (t). After an introduction of the state vector z = [ \u03b1 z \u03b1\u0307 z\u0307 ]T , the equations of motion can be derived by using the Lagrange formalism with the derivative z\u0307 = [ \u03b1\u0307 z\u0307 \u03b1\u0308 z\u0308 ]T of the state vector and \u03b1\u0308 = 2g sin(\u03b1) (M +m)\u2212ma\u03b1\u03072 sin(2\u03b1) + 2 cos(\u03b1)FC a (2M +m (1\u2212 cos(2\u03b1))) (31) as well as z\u0308 = mg sin(2\u03b1)\u2212 2ma sin(\u03b1)\u03b1\u03072 + 2FC 2M +m (1\u2212 cos(2\u03b1)) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003102_fpmc2016-1702-Figure19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003102_fpmc2016-1702-Figure19-1.png",
        "caption": "Fig. 19 Experimental device",
        "texts": [
            " Therefore, the authors removed the container and installed the heater inside the PARM to reduce the size of the actuator. Configuration of new actuator Fig. 17 shows photographs of the previous actuator and heater. Four constantan heaters (resistance: 3.7 \u2126) were connected in parallel and placed in the fixed container in which the GLPC occurred. Pressure was generated in the container and transmitted to the PARM to drive the actuator. Photographs of the miniaturized actuator and heater are shown in Fig. 18, and Fig. 19 shows an illustration of the combined actuator and heater. The fixed container was removed, and a heater of 9.3 \u2126 was directly installed inside the PARM. A pressure sensor (SMC PSE510-R06) was installed at the top of the PARM, and the device was filled with the working fluid. The weight of the working fluid was 3.2 g, and the total volume of the device is 1.87 cm3. Fig. 17 Previous actuator and heater 6 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess",
            " The device required 0.41 s to raise the pressure in the PARM from 0.3 to 0.35 MPa and 1.7 s to lower it again. The derivative of the pressure was 0.14 MPa/s right after the power was supplied and \u20130.06 MPa/s after the power was cut off. Comparing Figs. 20 and 7, the miniaturized actuator maintained the same level of pressure responsiveness despite the power supply being reduced to approximately 1/7 that of the previous actuator. An internal pressure control experiment was conducted using the device shown in Fig. 19. In this experiment, the parameters of the control system were set to the same values as in the previous experiment. The proportional gain and the integral gain were 400 V/Pa and 3 V/(Pa\u00b7s), respectively, and the reference pressure was set to increase from 0.3 to 0.35 MPa and then decrease back to 0.3 MPa. The configuration of the entire device is shown in Fig. 21, and a block diagram of the device is shown in Fig. 9. The data logger measured the pressure P and the voltage E of the heater during the experiment"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.10-1.png",
        "caption": "Fig. 6.10 (a) Milling machine. (b) Sample milling cutters and the resultant cut profiles",
        "texts": [
            " The shank length depends on how deep the hole is required to be. The lip angle depends on the level of physical support of the tool cutting edge. A larger lip angle supports cutting more aggressively under the same amount of point pressure, whereas a smaller lip angle cuts less aggressively. are its high material removal rate, high surface finishing, and dimensional control. The orientation of milling can be either vertical or horizontal. For instance, for vertical milling, a vertical milling machine with a vertical spindle axis is used (Fig. 6.10a). Some important features of milling can be modified via the following ways. By adjusting the power supply, the revolutions per minute of the milling cutter, which is held in the spindle, can be adjusted. Also, by applying different types of milling cutters (Fig. 6.10b), we can produce products with different milling properties. A selected milling cutter is mounted on the quill of a milling machine. The quill moves vertically in the head with an adjustable depth and contains the spindle installed with cutting tools. The quill feed handwheel moves the quill up and down within the head as does the quill feed lever. The knee moves up and down by sliding parallel to the column. The main function of the saddle on the knee is to allow for the translation of the worktable, and the column on the base holds the turret, which allows for the rotation of the milling head around the column\u2019s center"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002998_978-981-10-2404-7_30-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002998_978-981-10-2404-7_30-Figure1-1.png",
        "caption": "Fig. 1 Schematic diagram of system",
        "texts": [
            " studied the dynamic stability of biped robot in motion using the method of Lyapunov exponent (Yang and Wu 2006). As a result, the method of Lyapunov exponent can be used as a quantitative analysis method for the dynamic stability of the quadrotor UAV at the stage of takeoff, landing, and yawing. Compared with the direct method of Lyapunov, this method of Lyapunov exponent can be constructed. Particularly, the quantitative relationship between the structural parameters and the dynamic stability can be established and analyzed (Stephen et al. 2012). Figure 1 illustrates the coordinate system for the quadrotor UAV E(X, Y, Z) and the geodetic coordinate system B(x, y, z). There are two hypotheses, that is, the quadrotor UAV is rigid body and four axes of propeller are perpendicular to the plane of the body. When four forces are equal (F1 = F2 = F3 = F4), the quadrotor UAV is at the state of rising, falling, or hovering. The system is at the state of pitching,whenF2 = F4 and F1 6\u00bc F3. The system is at the state of rolling, when F1 = F3 and F2 6\u00bc F4. When F1 = F3 6\u00bc F2 = F4,the system is at the state of yawing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001244_esda2014-20131-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001244_esda2014-20131-Figure2-1.png",
        "caption": "Fig. 2 Schematic of dimensionless parameter on rotorbearing system",
        "texts": [
            " 4 2h Re     (4) The inertia effect is negligible if modified Reynolds number R\u0207<1 [15], so we may assumed the inertia effect to be neglected for low value of angular velocity. (3) Gas viscosity is not significantly changes in pressure and the temperature, so we may assume the gas viscosity to be constant. In Eq. 3, p is a gas pressure,  is viscosity, R 0 is gas constant, T 0 is absolute temperature and U is circumferential speed of rotor surfaces and h is the film thickness which is defined as h c ecos( )   (5) This Reynold\u2019s equation could be written in dimensionless form as in Eq. 6 using dimensional parameters (see Fig. 2); a 2 2 a a p P / P , h cH, x R , z R , 6 R 12 R U r , , P c P c                         3 3 0 s P P H P H P P H (P) (PH) (PH) t                          (6) where Pa is the atmospheric pressure, H is the dimensionless film thickness, c is the radial clearance and R is the radius of bearing. In the dimensionless Reynold\u2019s equation (Eq. 6) dimensional feeding parameter, orifice function and dimensionless bearing gap are defined in Eq. 7, Eq. 8 and Eq. 9 respectively; 2 0 0 0 0 2 a 12 d R T p c      (7) 2 k 1 k k d d d p u u u 11 k 12 d p u p p p2k for 1 r k 1 p p p (P) p2k 2 for r k 1 k 1 p                               (8) 2 2 x y 1 H( ) 1 cos( ), e e c       (9) where d0 is the orifice hole diameter, rp is the critical pressure ratio, k is the specific heat ratio, pd is pressure of orifice exhaust and pu is pressure of orifice inlet"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001171_amm.658.339-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001171_amm.658.339-Figure1-1.png",
        "caption": "Fig. 1. The design of the new hybrid ball bearing of the type 8S-4C.",
        "texts": [
            " To avoid the misalignment, we replaced in 7206C steel bearing just four All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 130.194.20.173, Monash University Library, Clayton, Australia-12/04/15,22:38:32) equidistant steel balls with silicon nitride balls of the same diameter, obtaining a new hybrid bearing with eight steel balls and 4 ceramic balls, denoted hereafter as 8S-4C ball bearing (Fig. 1). The results presented in literature [5] are obtained only for partial hybrid ball bearings functioning in contaminated lubricant with pure radial load. The effect of the replacement of steel balls with silicon nitride balls was observed just by post-experiment examination of the damaged surfaces of the raceways. In this paper tandems of face-to-face mounted 8S-4C bearings are tested at high speed, for uncontaminated mineral oil mist lubrication and pure axial load. To improve the accuracy of the measurements, a new testing device is proposed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002041_detc2015-46758-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002041_detc2015-46758-Figure1-1.png",
        "caption": "FIGURE 1. PRINCIPAL FIBER BUNDLE.",
        "texts": [
            " A curve \u0393 : I \u2282 R\u2192 P transversal to each fiber of P can be interpreted as a section in a one-dimensional set U , which is the image of the curve \u03b3 : I \u2282 R\u2192U . In a local trivialization, such curve is of the form (\u03b3(t),g(t)), where g : I \u2282 R\u2192 G is a curve in the group G, and \u03c0 \u25e6\u0393(t) = \u03b3(t). If G is a matrix group, the columns of the matrix are vectors in a base of a representation n-vector space V , so each curve in P has associated a base of V , which is not necessarily a frame in the tangent space of B (see Fig. 1). Given a principal fiber bundle (P,B,\u03c0,G) and g the corresponding Lie algebra of the Lie group G, the action G\u00d7P\u2192 P induces a homomorphism \u03d2 : g\u2192X (P) in the vector space of tangent vector fields in P. For each X \u2208 g the corresponding vector field X\u2020 = \u03d2(x) is known as a fundamental vector field, and is always vertical (that is, tangent to the fibers of the principal fiber bundle). Of course this vector space is also a Lie algebra and there is a canonical homomorphism with g. This vector field also satisfy (Rg)\u2217X \u2020 = \u03d2( ad(g\u22121)x), and the field X\u2020 = \u03d2(x), when restricted to a fiber, is the left invariant vector field of G"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001630_j.proeng.2014.03.133-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001630_j.proeng.2014.03.133-Figure1-1.png",
        "caption": "Fig. 1. Multilayer elastomeric structures examples: a) flat of rectangular shape, b) flat ring, c) cylindrical, d) conical, e) spherical.",
        "texts": [
            " Packages of thin-layered rubber-metal elements (TRME) successfully replace traditional technical systems, such as bearing, joints, compensating devices, shock-absorbers because of its important advantages: improving of machines dynamics, vibration and noise reducing, low shear and compression stiffness ratio [4-8]. These structures are used in machine building, shipbuilding, civil engineering, aviation and aerospace due to its unique mechanical properties. In practice the TRME packages of different geometric shapes are used: flat, cylindrical, conical and others; number of layers may be different, at least three (Fig. 1). In many applications of TRME structures it is necessary to know its stiffness characteristics, in particular, if TRME packet is used for vibration isolation of the object from vibrating base. The elastic compensation device mounted between the vibrating base and protected object is the main element of any passive vibration protection system. In this case the amplitude of the protected body oscillations depends on the excitation frequency and on the possibility of resonance phenomenon occurrence"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000976_20140313-3-in-3024.00064-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000976_20140313-3-in-3024.00064-Figure1-1.png",
        "caption": "Figure 1 : Rigid vehicle control parameters",
        "texts": [
            "00064 discussed. The autopilot design for the model using LQR technique is also discussed. In section 6, a comparison of the methods for autopilot design is made. Also, a comparison of the approximated model and integrated model is made. Finally, concluding remarks and directions for further works are given in the final section. The system modelling of RLV is done by analysing the forces and moments acting along a particular plane [3]. In this paper, the pitch plane is considered for system modelling. The figure 1 shows the control parameters along the pitch plane. Bending modes represent the flexibility of RLV at very high speeds. They couple with elements of the vehicle\u2019s control system and causes adverse effects upon vehicle loading and stability. So the effect of bending modes are also considered while formulating the system dynamics [1]. The equation of motion along the pitch plane is F m valong the path     L D mg T sin  (1) Where m is the reduced mass of the vehicle, v is the velocity along the longitudinal axis, L is the control engine thrust, D is the drag, T is the propellant pendulum angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003109_iciea.2016.7603844-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003109_iciea.2016.7603844-Figure4-1.png",
        "caption": "Fig. 4. Part of the motor temperature distribution diagram.",
        "texts": [
            " Motor main parameters PARAMETER Number of stator slots Number of rotor poles Rated Speed Peak Speed Rated Voltage Peak Power Peak Current Magnet residual induction(Br) at 25 Temperature coefficient of Br 24 16 1250rpm 5500rpm 245V 20KW 260A 1.26T -0.13%/ 1616978-1-4673-8644-9/16/$31.00 c\u00a92016 IEEE Fig. 2 shows the average eddy current loss in a magnet with different output power when the speed is 1250 rpm. When the output power is 15kw, the magnet loss is about 77.26w. Fig. 3 shows the temperature of the magnet in 15kw, the temperature can reach 175 . The motor temperature distribution diagram is shown in Fig. 4. The temperature already exceeds the permissible operation temperature of the permanent magnet. These are three reasons that generate the eddy current loss in the permanent magnet. That is slot harmonics, spatial harmonics and temporal harmonics of MMF. A. Slot harmonics The radial magnetic flux density (Br) in the air-gap through a slot in no-load is shown in Fig. 5. Its Fourier series are shown in Fig. 6. The radial magnetic flux density is lower under the stator slot and higher under the teeth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001883_055035-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001883_055035-Figure2-1.png",
        "caption": "Figure 2. An element of thin homogenous rod in bending.",
        "texts": [
            " Let the mass element be \u03c1 \u03c4=m xd ( )d with the volume element \u03c4 = s x xd ( )d , one obtains the rotational inertia to the oy-axis as \u03c1 \u03c4=I h x xd ( ) 12 ( )d . (2)yy When the rod deforms transversely, its every cross-section should produce shearing forces. Let the shearing force on the left of volume element be Q x t( , ) (down direction), and the right one be \u2032 = + \u2202Q Q x t Q x t x( , ) ( , )dx (up direction) with \u2202x being an abbreviation of \u2202 \u2202x. These two shearing forces form a force couple, which bends the rod. Figure 2 shows an element of a thin homogenous rod in bending. We consider the characteristic quantity of the transverse vibrations of a thin loaded rod, the displacement of the rod away from the equilibrium position along the z-direction at spacetime point (x,t), is u x t( , ). The curvature radius of the bending is = + \u2202 \u2202\u23a1\u23a3 \u23a4\u23a6( )R u u1 . (3)x xx 2 3 2 2 When the rod bends, the central line length dx remains unchanged. However, the upper part of the central line that suffers the tension of the nearby elements is prolonged; the lower part that suffers the pressure is compressed. Consequently, the force couple consists of tension and pressure, the so-called bending moment. Let the bending moment on the left of the volume element be M x t( , ) (clockwise), and the bending moment on the right be \u2032 = + \u2202M M x t M x t x( , ) ( , )dx (counter-clockwise). The bending moments act as resistance to the bending of the rod, and lead the system to the dynamic equilibrium states. Figure 3 shows the force acting on an element of a thin homogenous rod. As shown in figure 2, we take a lamina with thickness dz at z position and width w(x). We recall that the length is dx at the center line of the volume element. The length of the lamina is \u03b8+ = +R z x z x R( )d d d and the relative extension is z R. So that the tensile stress is = \u2212P Yz R with Y being Young\u2019s modulus. The tension element is dG = Pwdz, the bending moment element is = = \u2212M z G Yz w z Rd d d2 , and the bending moment is = \u2212M YJ R, (4) with \u222b= =J z w z s x r xd ( ) ( )2 2 the inertia moment per mass for the cross-section s(x) to the center \u2010oy axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003315_iecon.2016.7793383-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003315_iecon.2016.7793383-Figure16-1.png",
        "caption": "Fig. 16. Geometrical parameters of the sleeve",
        "texts": [
            " Therefore, it is considered that the lifetime will be increasing. From these results, it is confirmed that the airbag of SPET2-PE2-Kvr and Clean Cut are contributed to increase the lifetime of the RLAM. From these lifetime evaluation tests, we confirmed the effectiveness of the proposed airbag and Clean Cut sleeve to increase lifetime of RLAM. Here, summarized about the optimal diameter of the airbag that is derived from the geometric relationship of the sleeve. Geometry of the sleeve is shown as Fig. 16. In Fig. 16, L is the effective length, D is the diameter, q is the angle of a mesh of the sleeve, l is the length of the thread, and n is the number of turns of the thread around the sleeve. For using these parameters, the mechanical equilibrium model of the RLAM [10]is shown as follows: F = l n 1 4\u03c0n 3cos2\u03b8 \u22121( )\u2212 \u00b5 sin\u03b8 cos\u03b8( ) \u23a7 \u23a8 \u23a9 \u23ab \u23ac \u23ad P (1) l = L0 / cosq (2) n = l sinq / (pD) (3) Where, F is the contraction force, P is the pressure. \u00b5 is the friction coefficient between the sleeve and the airbag, the subscript 0 means the parameter when the RLAM is the natural length"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001378_s00707-015-1312-8-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001378_s00707-015-1312-8-Figure2-1.png",
        "caption": "Fig. 2 Planar separation of a beam",
        "texts": [
            " , N . (61) This type of separation is named \u2018spontaneous\u2019 by Landau and Lifshitz [32] and is applied by Cveticanin [31]. The main advantage of the suggested analytical procedure is its simplicity for practical use in comparison with the classical method based on the general principles of dynamics which have the vector form. Namely Eq. (34) means the use of the abstract and concise concept of generalized velocities to describe the problem via a simple equation. Let us consider a homogenous beam (Fig. 2) with mass M = \u03c1L A (62) and moment of inertia IS = 1 12 \u03c1AL3, (63) where \u03c1 is the density, L is the length, A = hs = const. is the constant cross section, h L is the width of beam and s L is the thickness of the beam. The beam moves planar in the x Oy plane and has three degrees of freedom. The generalized velocities of motion are the components of mass center velocity x\u0307S, y\u0307S and the angular velocity of the beam \u03c8\u0307. The kinetic energy of the beam is given with (36) Tb = 1 2 1 12 \u03c1AL3\u03c8\u03072 + 1 2 \u03c1L A ( x\u03072 S + y\u03072 S ) ",
            " Using the position coordinates of S1 xS1 = xS \u2212 \u03c1S1 cos\u03c8, yS1 = yS \u2212 \u03c1S1 sin\u03c8, (74) and \u03c1S1 = l/2, the velocity projections and the angular velocity are in general x\u0307S1 = x\u0307S + l 2 \u03c8\u0307 sin\u03c8 + x\u0307\u2217 S1, y\u0307S1 = y\u0307S \u2212 l 2 \u03c8\u0307 cos\u03c8 + y\u0307\u2217 S1, \u03c8\u03071 = \u03c8\u0307 + \u03c8\u0307\u2217 1 , (75) and x\u0307\u2217 S1, y\u0307\u2217 S1 and \u03c8\u0307\u2217 1 are velocity and angular velocity perturbation values. Substituting (75) into (73) we have Ta1 = 1 2 1 12 \u03c1A(L \u2212 l)3 ( \u03c8\u0307 + \u03c8\u0307\u2217 1 )2 +1 2 \u03c1(L \u2212 l)A (( x\u0307S + l 2 \u03c8\u0307 sin\u03c8 + x\u0307\u2217 S1 )2 + ( y\u0307S \u2212 l 2 \u03c8\u0307 cos\u03c8 + y\u0307\u2217 S1 )2 ) . (76) For the impulse shown in Fig. 2, it is J = \u2212J sin(\u03c8 + \u03b1)i + J cos(\u03c8 + \u03b1)j, (77) where \u03b1 = const. and J is the intensity of impulse of the impact force. As the impulse acts on the end of the beam (see Fig. 2) and its component normal on the beam has the same direction as the angular velocity of the beam, the virtual work of the impulse is according to (50) \u03b4A = \u2212J sin(\u03c8 + \u03b1)\u03b4 x\u0307S + J cos(\u03c8 + \u03b1)\u03b4 y\u0307S + J ( L 2 cos\u03b1\u03b4\u03c8\u0307 ) . (78) Due to (53) and (78), the generalized impulses are Qx\u0307 S = \u2202A \u2202 x\u0307S = \u2212J sin(\u03c8 + \u03b1), Qy\u0307S = \u2202A \u2202 y\u0307S = J cos(\u03c8 + \u03b1), Q\u03c8\u0307 = \u2202A \u2202\u03c8\u0307 = J L 2 cos\u03b1. (79) Using the relations (34) \u2202 \u2202 x\u0307S (Ta1 + Ta2 \u2212 Tb) = Qx\u0307 S, \u2202 \u2202 y\u0307S (Ta1 + Ta2 \u2212 Tb) = Qy\u0307S, \u2202 \u2202\u03c8\u0307 (Ta1 + Ta2 \u2212 Tb) = Q\u03c8\u0307S, (80) the velocity and angular velocity of the remainder beam are obtained"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.9-1.png",
        "caption": "Fig. 3.9 a Schematic representation of the normal stress distribution \u03c3x = \u03c3x(z) of a bending beam; bDefinition and position of an infinitesimal surface element for the derivation of the resulting moment action resulting from the normal stress distribution",
        "texts": [
            " 11Note that according to the assumptions of the Bernoulli beam the lengths 01 and 0\u20321\u2032 remain unchanged. 3.2 Derivation of the Governing Differential Equation 97 3.2.2 Constitutive Equation The one-dimensionalHooke\u2019s law according to Eq. (2.3) can also be assumed in the case of the bending beam, since, according to the requirement, only normal stresses are regarded in this section: \u03c3x = E\u03b5x. (3.20) Through thekinematics relation according toEq. (3.16), the stress results as a function of deflection to: \u03c3x(x, z) = \u2212Ez d2uz(x) dx2 . (3.21) The stress distribution shown in Fig. 3.9a generates the internal moment, which acts in this cross section. To calculate this internal moment, the stress is multiplied by a surface, so that the resulting force is obtained. Multiplication with the corresponding lever arm then gives the internal moment. Since the stress is linearly distributed over the height, the evaluation is done for an infinitesimally small surface element: dMy = (+z)(+\u03c3x)dA = z\u03c3xdA. (3.22) 98 3 Euler\u2013Bernoulli Beams and Frames Therefore, the entire moment results via integration over the entire surface in: My = \u222b A z\u03c3xdA (3",
            " In the case of plane bending with My(x) = const., the bending line can be approximated in each case locally through a circle of curvature, see Fig. 3.10. Therefore, the result for pure bending according to Eq. (3.25) can be transferred to the case of plane bending as: \u2212 EIy d2uz(x) dx2 = My(x). (3.27) Let us note at the end of this section that Hooke\u2019s law in the form of Eq. (3.20) is not so easy to apply12 in the case of beams since the stress and strain is linearly changing over the height of the cross section, see Eq. (3.26) and Fig. 3.9. Thus, it might be easier to apply a so-called stress resultant or generalized stress, i.e. a simplified representation of the normal stress state13 based on the acting bending moment: My(x) = \u222b\u222b z\u03c3x(x, z) dA, (3.28) which was already introduced in Eq. (3.22). Using in addition the curvature14 \u03ba = \u03ba(x) (see Eq. (3.16)) instead of the strain \u03b5x = \u03b5x(x, z), the constitutive equation can be easier expressed as shown in Fig. 3.11. The variables My and \u03ba have both the advantage that they are constant for any location x of the beam",
            "2 Derivation of the Governing Differential Equation 195 dQz(x) dx = \u2212qz(x), (4.19) dMy(x) dx = +Qz(x). (4.20) 4.2.3 Constitutive Equation For the consideration of the constitutive relation,Hooke\u2019s law for a one-dimensional normal stress state and for a one-dimensional shear stress state is used: \u03c3x = E\u03b5x , (4.21) \u03c4xz = G\u03b3xz, (4.22) whereupon the shear modulus G can be calculated based on the Young\u2019s modulus E and the Poisson\u2019s ratio \u03bd as: G = E 2(1 + \u03bd) . (4.23) According to the equilibrium configuration of Fig. 3.9 and Eq. (3.22), the relation between the internal moment and the bending stress can be used for the Timoshenko beam as follows: dMy = (+z)(+\u03c3x )dA, (4.24) or alternatively after integration under the consideration of the constitutive equation (4.21) and the kinematics relation (4.17): My(x) = +E Iy d\u03c6y(x) dx . (4.25) The relation between shear force and cross-sectional rotation results from the equilibrium equation (4.20) as: Qz(x) = +dMy(x) dx = +E Iy d2\u03c6y(x) dx2 . (4.26) Before looking in more detail at the differential equations of the bending line, let us summarize the basic equations for the Timoshenko beam in Table4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.41-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.41-1.png",
        "caption": "FIGURE 8.41",
        "texts": [
            " Note that these sections are translated and properly oriented to compose an overall waterslide configuration by the following translation and rotation operations: X\u00f0u;w\u00de \u00bc X0 \u00fe T\u00f0q\u00dex\u00f0u;w\u00de \u00bc 2 64 X01 X02 X03 3 75\u00fe 2 64 cos q 0 sin q 0 1 0 sin q 0 cos q 3 75 2 64 x1\u00f0u;w\u00de x2\u00f0u;w\u00de x3\u00f0u;w\u00de 3 75 (8.112) where T(q) is the rotational matrix that orients the section by rotating through an angle q about the X2 (or Y) axis. X0i is the location of the local coordinate system of the section in the waterslide configuration. x(u,w) is the surface function of the flume section referring to its local coordinate system. As discussed in Section 8.2.1, the Lagrange equation of motion based on Hamilton\u2019s principle (Kane, 1985) for this particle dynamic problem, shown in Figure 8.41, can be stated as d dt vL v _q vL vq \u00bc Q (8.113) where The Lagrangian function L is defined as L h T V, _q \u00bc vq=vt. The generalized coordinates q in this waterslide application are the parametric coordinates of the surface (i.e., q \u00bc [u, w]T). Object path and unit vectors for friction forces on a straight flume section. When the system is conservative, Q \u00bc 0. For a nonconservative system, Q \u00bc F, where F is the vector of generalized friction forces. For this motion analysis problem, the kinetic energy T and the potential energy U are, respectively, T \u00bc m 2 _X u;w 2 and U \u00bc mgX2\u00f0u;w\u00de (8.114) where m is the particle mass and g is the gravitational acceleration. For friction cases, the generalized friction forces Q \u00bc [fu, fw] T can be derived as (Chang, 2007) fu \u00bc m\u00f0g\u00fe an\u00de$n\u00f0et$X;u\u00de; and fw \u00bc m\u00f0g\u00fe an\u00de$n\u00f0et$X;w\u00de (8.115) where m is the friction coefficient. n is the unit normal surface vector (shown in Figure 8.41). an is the normal acceleration of the riding object. et is the unit vector along the tangential direction of the object\u2019s path, which is also the direction of the object\u2019s velocity _X and the tangential acceleration at. Following Eq. 8.113, two coupled second-order ordinary differential equations that govern the particle motion can be obtained as k0\u20acu \u00bc k1 _u 2 \u00fe k2 _w2 \u00fe k3 _u _w\u00fe k4 (8.116a) k0 \u20acw \u00bc k5 _u 2 \u00fe k6 _w2 \u00fe k7 _u _w\u00fe k8 (8.116b) where k0 through k8 consist of polynomials of u and w and their products"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002884_1.4034511-Figure18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002884_1.4034511-Figure18-1.png",
        "caption": "Fig. 18 Bearing mock-ups with heat dissipation model",
        "texts": [
            " From the figure, it is noted that the temperature drop profile across the interface with lubrication is very less compared to dry condition irrespective of interface temperature, because of the improved thermal conductivity of the grease which decreases the gap resistance considerably. 5.3.3 Simulation of LMMB. Further, an attempt is made to simulate the temperature distribution of LMMB with varying number of balls at constant surface roughness (2.08 lm) and interface temperature (385 K) under dry and lubricating conditions. These works have been taken to study the temperature dissipation of the LMMB by varying the number of balls with available TCR data obtained from both with and without lubrication, and the FEM model is shown in Fig. 18. From the above simulation results, it is observed that the temperature distribution varies with varying number of balls under dry condition. The same phenomenon was observed under lubricating condition. It can be seen that Table 1 List of uncertainty sources and results for TCR S. no. Variables Measured values Reference value Deviations Percentage of error 1 Thermocouples ( C) 42.57 43 0.043 1.00 2 Thermocouples locations/fixation ( C) 46.95 50 3.05 06.10 3 Insulation of stack system (W/m2) 2.89 3 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.6-1.png",
        "caption": "Figure 9.6 Space-vector diagram of a nonsalient-pole PMSM in pu values. Ld= Lq= 0.5. There is a stator reference frame (xy) and a rotor reference frame (dq). The machine operates as a motor at its nominal operating point in light field weakening (small negative id= 0.24). \u03c9s = 1, us= 1, is= 1, and \u03c6 15\u00b0. The load angle \u03b4s 29\u00b0. Power factor and torque have equal values, that is, cos\u03c6s 0.965 and Te= 0.965. \u03b3 = 104\u00b0 is the electric current angle measured from the d-axis.",
        "texts": [
            " The machine is operating in light field weakening, that is, id= 0.24 pu in the negative direction. Resistance is neglected, the pu inductances are Ld= Lq= 0.5, and rotor position angle is \u03b8r= 40\u00b0 measured from the x- to d-axis, in which \u03c8PM is located. Determine the load angle \u03b4s, the PM linkage flux \u03c8PM, the d- and q-components of current and voltage, the power factor cos \u03c6, the electric current angle \u03b3 measured from the d-axis to the stator current space vector, and the electromagnetic torque Te. SOLUTION: Figure 9.6 shows the space-vector diagram. There is a stator reference frame xy and rotor reference frame dq shifted by rotor angle \u03b8r from the x-axis. The PM linkage flux \u03c8PM, the value of which will be determined later, is along the d-axis. The stator flux linkage \u03c8 s is shifted from the d-axis by a load angle that will now be calculated. Applying the Pythagorean theorem, \u03c8d can be written as follows. 2 12\u03c8d \u03c8 scos \u03b4s \u03c82 Lqiq 0.5 0.97 2 0.874pus where i2 i2 12 0.242 0.97puiq s d Since \u03c9s= 1 pu, the cosine function can be calculated to determine the load angle",
            " The torque equation is used as a development starting point for the various PM machine control principles. In stator current control, using electric current angle \u03b3 is easier than using the load angle \u03b4s, because electric current angle is a vector control parameter. Section 9.5 will show load angle \u03b4s to be a direct torque control (DTC) parameter. Per-unit power can be expressed pu voltage and current as follows. P usiscos \u03c6 (9.17) The power-factor angle \u03c6 can be written in terms of \u03b3 and \u03b4s on the basis of Figure 9.6 to \u03c0become \u03b3 . Therefore\u03b4s2 \u03c0 P usiscos \u03b3 (9.18)\u03b4s2 The pu value of power is as follows. \u03c0 P usiscos 2 \u03b4s \u03b3 usissin \u03b3 \u03b4s (9.19) Equation (9.19) can also be written P usissin \u03b3 \u03b4s usis sin \u03b3 cos \u03b4s cos \u03b3 sin \u03b4s (9.20) Figure 9.6 reveals that id iscos \u03b3 (9.21) iq issin \u03b3 (9.22) ud ussin \u03b4s \u03c9sLqiq (9.23) uq uscos \u03b4s \u03c9s \u03c8PM Ldid (9.24) The pu power can be expressed P \u03c9s \u03c8PMissin \u03b3 i2sin 2\u03b3s Lq 2 Ld (9.25) Correspondingly, the torque becomes P Lq LdTe is \u03c8PM sin \u03b3 i2sin 2\u03b3 (9.26)s 2\u03c9s A so-called characteristic current ix, which differs slightly from iPM, has been introduced and is preferred by some researchers, for example, Soong et al. 2007a and 2007b. \u03c8PMix (9.27) Ld This value of this characteristic current determines one aspect of the motor drive\u2019s nature",
            " In so doing, power factor begins to increase to its maximum at cos\u03c6= 1 before going capacitive at higher speeds. This behaviour is illustrated in Figure 9.11 for a rotor surface magnet PMSM with pu values of Ls= 0.6, \u03c8PM= 1, is= 1, and us= 1. EXAMPLE 9.4: Produce a diagram Te, P, cos\u03c6= f (\u03c9s) that shows the field weakening dependence for a rotor surface magnet PMSM with Ld=Lq=Ls= 0.6, \u03c8PM= 1, is= 1, and us= 1. SOLUTION: The space-vector diagram is similar to that of Figures 9.6 and 9.10, however, in Figure 9.6, the different inductance results in changes in value for stator linkage flux and other parameters. For an id= 0 control, iq= is= 1. The voltage increases to the nominal value when the pu value us is 1. 2 12\u03c82 0.6 1 2 1.16Lsiq\u03c8 s PM 1\u03c8PMcos \u03b4s 0.86 \u03b4s 30.68\u00b0 \u03c8 s 1.16 1 us \u2245 es \u03c9s\u03c8 1 \u03c9s 0.86s 1.16 With id= 0, is= iq= 1. In a nonsalient-pole machine using id= 0 control, the angle shift between the voltage and current is identical with \u03b4s= 30.7\u00b0, because the q-axis is perpendicular to the d-axis, and the voltage vector us is perpendicular to the stator linkage flux \u03c8s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure5-1.png",
        "caption": "Fig. 5 Tyre contact forces applied to the upright.",
        "texts": [
            " The caliper is rigidly connected to the upright through the two connection points. A tangential braking force is uniformly applied on the pads supports. This force is given by the friction generated when the pads are pressed against the disc surface. A uniform pressure load is applied in the six cylinders of the caliper to model the hydraulic pressure acting in the caliper. Loads acting on the upright come from the tyre/road contact forces. The lateral force is applied at the contact patch that is rigidly connected to the upright centre as shown in Fig. 5. The upright centre is rigidly connected to the bearing houses. The longitudinal and vertical forces Fx and Fz are applied at the upright centre. The resultant vertical force Fz is obtained through the following relation bvert FFFz (3) Where Fvert is the vertical tyre force acting at the contact patch, while Fb is the braking force that acts at the pads supports. The moment Mz is given by the product of the longitudinal force Fx and the steering arm, i.e. the distance between the centre of the contact patch and the centre of the upright in y direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure6-1.png",
        "caption": "Figure 6. The system of equivalent springs and dampers for the system shown in Figure 5(a).",
        "texts": [
            " the angle c (see equation (20)) tan 2 c \u00bc P3 j\u00bc1 cj sin 2 \u2019jP3 j\u00bc1 cj cos 2 \u2019j \u00bc 2c sin 10 =3\u00f0 \u00de \u00fe c sin 2 =3\u00f0 \u00de \u00fe 3c sin 2 \u00f0 \u00de 2c cos 10 =3\u00f0 \u00de \u00fe c cos 2 =3\u00f0 \u00de \u00fe 3c cos 2 \u00f0 \u00de \u00bc ffiffiffi 3 p 3 \u00f027\u00de the solution of which is c \u00bc 12 \u00bc 15o \u00f028\u00de Equation (21a,b) yields the values of two damping coefficients cI \u00bc X3 j\u00bc1 cj cos 2 \u2019j c \u00bc 2c cos2 5 =3\u00fe =12\u00f0 \u00de \u00fe c cos2 =3\u00fe =12\u00f0 \u00de \u00fe 3c cos2 \u00fe =12\u00f0 \u00de \u00bc 3:866 c, cII \u00bc X3 j\u00bc1 cj sin 2 \u2019j c \u00bc 2c sin2 5 =3\u00fe =12\u00f0 \u00de \u00fe c sin2 =3\u00fe =12\u00f0 \u00de \u00fe 3c sin2 \u00fe =12\u00f0 \u00de \u00bc 2:134 c \u00f029a; b\u00de These two dampers are shown in Figure 6. Let us find now the system of equivalent springs. The position of one of the spring follows from equation (11a) tan 2 k \u00bc P3 i\u00bc1 ki sin 2\u2019iP3 i\u00bc1 ki cos 2\u2019i \u00bc k sin 8 =3\u00f0 \u00de \u00fe 2k sin 4 =3\u00f0 \u00de \u00fe 3k sin 0\u00f0 \u00de k cos 8 =3\u00f0 \u00de \u00fe 2k cos 4 =3\u00f0 \u00de \u00fe 3k cos 0\u00f0 \u00de \u00bc ffiffiffi 3 p 3 \u00f030\u00de which gives k \u00bc 12 \u00bc 15o \u00f031\u00de Equations (11b,c) are then used to calculate the stiffness coefficients of the equivalent springs kI\u00bc X3 i\u00bc1 ki cos 2 \u2019i k\u00f0 \u00de\u00bckcos2 4 =3\u00fe =12\u00f0 \u00de \u00fe2kcos2 2 =3\u00fe =12\u00f0 \u00de\u00fe3kcos2 0\u00fe =12\u00f0 \u00de\u00bc3:866k, kII\u00bc X3 i\u00bc1 ki sin 2 \u2019i k\u00f0 \u00de\u00bcksin2 4 =3\u00fe =12\u00f0 \u00de \u00fe2ksin2 2 =3\u00fe =12\u00f0 \u00de\u00fe3ksin2 0\u00fe =12\u00f0 \u00de\u00bc2:134k \u00f032a\u00de These springs are shown in Figure 6. Note that kI \u00fe kII \u00bc 6k, as stated in Corollary 1.2 and equation (13). If there is no damping, the natural frequencies of free vibration are easily calculated as !1 \u00bc ffiffiffiffiffiffi kII m r 1:461 ffiffiffiffi k m r , !2 \u00bc ffiffiffiffi kI m r 1:966 ffiffiffiffi k m r \u00f033a; b\u00de at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from In order to emphasize some of the benefits of the approach presented earlier, let us show the alternative way of finding the natural frequencies",
            "2 \u00bc B A !\u00bc!2 \u00bc 15k 4m!2 2ffiffiffi 3 p k \u00bc 2\u00fe ffiffiffi 3 p \u00bc 0:268 \u00f041a;b\u00de The corresponding modes shapes are given in Figure 7(b), where the one corresponding to !1 is plotted as a dashed-dotted line and the one corresponding to !2 as a dotted line, and the value of A is taken arbitrarily to be equal to unity. It is important to note that the angles between these two directions and the horizontal are 75o and 15o, which agrees with the position of two-element equivalent system of springs shown in Figure 6. However, unlike this procedure related to natural modes of vibration, which involves lengthy calculations, the one related to the equivalent system of springs performed previously is straightforward and mathematically tractable. Example 2. The second example is shown and defined in Figure 8(a). The system consists of four springs, whose stiffness coefficients are k1\u00bc 4000N/m, k2\u00bc 2000N/m, k3\u00bc 2500N/m and k4\u00bc 1184.6N/m. They are joined in a point M0, where the particle of mass m (mg\u00bc 900N) is located"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002436_j.cirp.2015.04.129-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002436_j.cirp.2015.04.129-Figure4-1.png",
        "caption": "Fig. 4. Schematic illustration of forming operation. (a) Prior to forming and (b) after forming operation.",
        "texts": [
            " By using the top view, the position of the centre of the mference of the sphere is determined by fitting a 3-point-circle, ch is then compared to the position of the centre of the rod. The ntricity of the forming product is detected accordingly. As n, burr formation does not affect the measuring method for ntricity but the direction of burr formation is also detected. The form filling of the cavity is regarded with respect to forming upper radius ru, lower radius rb as well as the absence of any burr formation (Fig. 4b). The transition from lateral surface of the cavity to the cavity ground is not sharp but a radius rd due to manufacturing process. The radii ru and rb are measured using side views of the forming product. Each forming product is surveyed from two directions thus giving four measuring values for each radius. Due to measuring inaccuracy, radii 5 mm are defined as completely filled. The angle of burr formation is detected in top view. Again, measuring is done with Keyence VHX 1000. To improve legibility, the quotient of the volume of preform VP relative to the volume of the die VD, called relative volume VR, is introduced and defined as: VR \u00bc VP=VD; (2) well knowing that no generalizable correlation between burr formation or eccentricity and VR can be derived as VD is constant",
            " It can be derived that an increasing relative volume of preform requires decreasing values of eccentricity to avoid burr formation. For instance, at a relative volume of preform VR = 103%, eccentricity of more than 80 mm can be tolerated without burr formation. This value is reduced to 25 mm for VR = 127%. For VR 128% burr formation takes place for any preform in cavity \u2018\u2018A\u2019\u2019 and \u2018\u2018B\u2019\u2019 being formed. Generally, the volume of the preform needs to be larger than the volume of the cavity because there is no axial fixation of the shaft in the lower die (Fig. 4a) so that the shaft, and thus some fraction of the preform volume (between 4% and 30% of VP; proportional to VR), is forged through the lower die. In order to find show to ru \u2018\u2018D\u2019\u2019 cavi to th grin 4.4. T allow prod is se eith area cavi whi for V geom win cavi F fille in Fi and s little correlation with VR. Complete form filling with regard is achieved for VR 118%. The form filling of cavity \u2018\u2018C\u2019\u2019 and is not considered in Fig. 8. This is due to the geometry of the ties because the transition from the lateral surface of the cavity e cavity ground is not a constant value which is due to the fineding process"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002278_6.2015-3288-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002278_6.2015-3288-Figure11-1.png",
        "caption": "Figure 11: Power Module",
        "texts": [],
        "surrounding_texts": [
            "A power module has a 6-pos connector that provides +5.3V voltage and 2.25A current. It is needed as we need a connector to power the ArduPilot Mega. The power module accepts a maximum input voltage of 18V (up to 4S Lipo battery) and maximum current of 90Amps."
        ]
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.16-1.png",
        "caption": "FIGURE 8.16",
        "texts": [
            " With Adams/Car, users can simply enter vehicle model data into the templates and the program automatically constructs subsystem models (e.g., engine, shock absorbers, tires) as well as full vehicle assemblies. Once these templates are created, they can be made available to novice users of the software, enabling them to perform standardized vehicle maneuvers. A Formula SAE racecar model developed by engineering students was converted into an Adams/ Car model using the provided templates, as shown in Figure 8.16(a). The vehicle model was then simulated for skid pad racing, a constant-radius cornering simulation shown in Figure 8.16(b). In this section, four case studies and two tutorial examples are presented. The case studies involve a broad range of applications, including a kinematic study of a racecar suspension, the design of a HMMWV suspension, driving simulators, and recreational waterslides. The purpose of these case studies is to demonstrate the engineering capabilities of motion analysis software and some of its common industry applications. Tutorial examples, including a sliding block mechanism and a singlepiston engine, are also presented"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure1-1.png",
        "caption": "Figure 1 Schematic of axis layout of a skiving process.",
        "texts": [
            " Particularly, understanding the effects of cutter accuracy and of cutter set up deviation on the skived gear accuracy, and optimizing them are indispensable to ensure reliability in an actual gear production. Unfortunately, few studies on those issues can be found. In this paper, geometrical model that can predict the effect of pitch deviation and of run out of a cutter on skived gear accuracy is proposed. Experiments were also carried out to verify the validity of the model, and the results were in good agreement simulated results. 2 DESCRIPTION OF SKIVING As shown in Fig. 1, in a skiving, shaft angle between the cutter and the workpiece axis is given, the circumferential velocity vectors of the cutter and workpiece have different directions. As a result, a sliding velocity vector is generated and it serves as a cutting action that removes the stock as a chip. Table 1 shows the internal gear data, and the Figure 2 shows a photograph of the skived gear. Table 1 Gear data Table 2 Cutter data Parameter Value Module 1.6 Number of teeth 88 Pressure angle 20o Helix angle 18o (LH) Base circle diameter 148"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000672_978081000342.375-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000672_978081000342.375-Figure4-1.png",
        "caption": "Figure 4: Schematic design and rotor model of the turbocharger",
        "texts": [
            " Though, some experimental parameters such as oil supply temperature and the mass moments of inertia of the disk are not defined exactly in [16]. In the prediction the support structure is assumed as ideally rigid and an interaction with the rotor-bearing system is neglected. The influence of the drive is also not taken into account. Consequently, an exact agreement between measurement and prediction cannot be expected. Therefore, the good correspondence between the predicted and measured phenomena can be described as very satisfactorily. Figure 2: Overhung rotor model Figure 4 shows a typical design of a bearing setup for small automotive turbochargers including two FRB and a double acting thrust bearing with two collars. It is similar to the one investigated in this paper. Further, Figure 4 includes the discretized rotor model. The rotor consists of a flexible shaft with added masses. Whereas the floating ring bearings are positioned at node seven and eight the thrust bearing is added on node five. Compressor and turbine wheel are arranged on nodes four and eleven. It is assumed that unbalances exist on the turbine and compressor wheel with a phase shift of 180\u00b0. The characteristic parameters of the rotor and the bearings are listed in Table 1. Table 1: Characteristic parameters Parameter Value Total rotor length lR in mm \u2248 100 Total rotor mass mR in g \u2248 100 Lubricant 5W40 Oil supply temperature Tsup in \u00b0C 90 FRB length Bi / Ba in mm 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.7-1.png",
        "caption": "Fig. 3.7 Example for spatial kinematic chains: five-point wheel suspension",
        "texts": [
            " The bodies have three rotational motion components and no translational components. The relative motions of the bodies can only rotate around the axes going through the fixed point. Spatial kinematic chains In spatial kinematic chains, the motion of the bodies can be described as general spatial. In general there are three translational and three rotational motion components. The relative motion of the bodies in the joints\u2014depending on the joint\u2014is as well arbitrary. This is illustrated with the complex example of a five-point wheel suspension in Fig. 3.7. A joint is defined as a connection between two adjacent bodies within a kinematic chain. Depending on the degrees of freedom fGi of a joint, it generates 6 fGi geometric constraints between the two adjacent bodies. The relative motion between two bodies in a kinematic chain which are connected by a joint can be described by so-called natural or, respectively, relative 0 coordinates bi. In Tables 3.1 and 3.2, useful joints for a large class of technically applicable cases are presented. A detailed analysis of the corresponding spatial kinematic chains will be given in subsequent chapters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003077_urai.2016.7625772-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003077_urai.2016.7625772-Figure1-1.png",
        "caption": "Fig. 1. Planar cable driven parallel robot.",
        "texts": [
            "eywords - Cable robot, Simulation, Calibration, Laser distance sensor. 1. Introduction A cable-driven parallel robot, also simply called a cable robot [9, 10], is a parallel kinematic machine that mainly consists of an end-effector(EE), cables, winches, pulleys and rigid frame as shown in figure 1. The cables connect the EE to the winches that control the EE pose by changing the cable length. Cable robots allow high acceleration and high speed [I, 2] and high payloads [3] and large workspace [4,9]. The pulleys can be attached to structures such as steel frames or walls, which allows to change the robot configuration quickly. In such applications, calibration is one of the key tasks to improve the accuracy. When we have a new configuration that requires the identification of the actual geometric parameters, a simple method using cheap equipment is highly desirable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.33-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.33-1.png",
        "caption": "Fig. 2.33 Three-element truss structure with force boundary condition",
        "texts": [
            " u1X = u1Y = u3X = u3Y = 0, gives the following reduced system of equations: E A\u221a 2a \u23a1 \u23a3 1 0 0 1 \u23a4 \u23a6 \u23a1 \u23a3u2X u2Y \u23a4 \u23a6 = \u23a1 \u23a30 0 \u23a4 \u23a6 . (2.221) \u2022 All nodal displacements \u2013 u1X = u1Y = u3X = u3Y = 0, \u2013 u2X = u X , u2Y = uY . \u2022 Elemental forces in each rod General: \u03c3 = E L (\u2212u1 + u2) \u21d2 F = E A L (\u2212u1 + u2). Thus: F = E A L (\u2212 cos(\u03b1)u1X \u2212 sin(\u03b1)u1Y + cos(\u03b1)u2X + sin(\u03b1)u2Y ). Our case: FI = E A\u221a 2a ( +1 2 \u221a 2u X \u2212 1 2 \u221a 2uY ) = E A 2a (u X \u2212 uY ), (2.222) FII = E A\u221a 2a ( +1 2 \u221a 2u X + 1 2 \u221a 2uY ) = E A 2a (u X + uY ). (2.223) 2.7 Example: Plane truss structure with three rod elements The following Fig. 2.33 shows a two-dimensional truss structure. The three rod elements have the sameYoung\u2019smodulus E and length L . However, the cross-sectional areas Ai (i = I, II, III) are different from rod to rod. The structure is loaded by a point load F0 at node 2. Determine: \u2022 the free body diagram, \u2022 the global stiffness matrix, 2.4 Assembly of Elements to Plane Truss Structures 71 \u2022 the reduced system of equations under consideration of the boundary conditions, \u2022 the nodal displacements at node 2. \u2022 Simplify the nodal displacements at node 2 for the special case AI = AII = AIII = A"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.21-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.21-1.png",
        "caption": "Fig. 3.21 Kinematic transformers",
        "texts": [
            " The advantage of this method is that the global relations are clearly described (also easily interpretable), and the complete system of equations can often be explicitly solved. A disadvantage is that the initialization in the topological format can sometimes become complex. In the example of the double wishbone suspension, the system consists of the two coupled kinematic loops L1 and L2, where these loops are coupled by the input angle b1 of the upper trailing arm, as well as by the three angles b2; b3; b4 of the spherical joint S1 from Fig. 3.21. The entire system for the relative kinematics can be combined as follows: Example Double wishbone wheel suspension The system consists of 2 kinematic transmission elements (kinematic transformers) L1 and L2 3 linear coupling equations The DoF are determined by: 2 Kinematic loops, one with two DoF and the other with five DoF 7 3 Coupling equations -3 1 Branching -1 Degrees of freedom 3 (including one isolated DoF (Fig. 3.21)) The corresponding block diagram can be found in Fig. 3.21 and will be discussed in further detail in the next section. Notes \u2022 The kinematic analysis of a multibody system with kinematic loops, which is based on the topological method, will be analyzed using the example of the double wishbone axle in Sect. 3.5.5. \u2022 Besides the three principles for the formulation of the kinematic equations of complex multibody systems introduced in Sect. 3.4, there are further approaches, which, among other details, use particular properties of the topological structure of the system, leading to an implicit structure of the constraint equation system, which can be numerically solved",
            " The four-link mechanism can from now on be considered as a kinematic transformer with a non-linear input-output behavior. It transforms the input angle b into the output angle c. This concept can be expanded further, as shown in Sect. 3.5.4. After the coupling equations between different kinematic transformers are determined, the individual kinematic transformers, which comprise the non-linear\u2013 but local\u2013transmission equations of the disconnected multibody loops can be assembled over a block diagram to a kinematic network (compare Fig. 3.21), whose detailed kinematic analysis will be developed in the next section. The global constraint equations, as the result of the global assembly of the mechanism in modular form, are now available for further steps of the analysis. The structure based on this approach is extremely advantageous for the handling of complex multibody systems with kinematic loops: Every module, i.e. every kinematic transformer, includes only a limited number of local nonlinear constraint equations, which are manageable and can be solved independently from other transformers in the system: the structure of these constraint equations is only dependent on the geometric parameters inside the local loop"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002933_gt2016-56900-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002933_gt2016-56900-Figure14-1.png",
        "caption": "Figure 14. The structure of back-to-back rotor",
        "texts": [
            " Siffness-parameters of AMB-HPS actuator The results of the calibration show that this AMB-HPS can provide a current stiffness up to 410N/A at the bias current Ib=5A, and this means that the control force reaches 410N when the perturbation current Ip reaches 1A. This force is big enough to do the vibration control for the rotor system. The AMB can exert a control force at any frequency as said previously, such as synchronous [19], multi-harmonic vibration [20]; this paper focuses on the solution for instability. To investigate the control strategy, a typical back-to-back compressor faced with rotordynamic instability was taken as example. Figure 14 shows the structure of the back-to-back rotor composed with three low pressure impellers and three high pressure impellers. The rotor is 2660 mm in length, the bearing span distance is 1987 mm, and the diameter of bearing journal is 150 mm. Two five-pad tilting pad bearings support the rotor. The width of the pad is 79.5 mm and radius clearance of bearing is 0.12 mm. The rated rotating speed of the compressor is 10700r/min. Figure15 indicates the rotordynamic model of the rotor modeled with 55 elements, and the bearings locate at 5th and 35th node",
            "url=/data/conferences/asmep/89517/ on 02/04/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use stability with the increasing preload. In this simulation, the preload is 0.2 to make the rotor more stable when the other parameters are same. In centrifugal compressor, nearly every seal and blade can cause the cross coupling stiffness, which will lead the rotor unstable. Here we assumed that force caused by the cross coupling stiffness exerted on the location of balance piston (this is the most serious condition). As Fig.14 indicates, the cross coupling force exerted on rotor and the unbalance force on the rotor are represented by Fkcr and Fu respectively. Equation (5) describes the cross coupling force that exerts at balance piston, and the cross stiffness q is assumed as 8e6 N/m that causes the rotor unstable when the rotating speed arrives at 9000r/min. The aim of this assumption is just to compare the effect of two types control strategy for the instability problem. Moreover, the excitation force is the unbalance force that described in Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.40-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.40-1.png",
        "caption": "FIGURE 8.40",
        "texts": [
            " Safety problems discovered after the slide is built and installed are usually too late and too costly to correct. In this study the riding object was assumed to be a particle with concentrated mass. Flume sections were represented in a CAD environment using geometric dimensions such as height and width. Friction forces between the riding object and the flume surface were also incorporated. Basic sections of the flume, such as straight, elbow, and curved, serve as the building blocks for composing waterslide configurations (see Figure 8.40). In addition, guard sections (essentially vertical walls) are added to reinforce safety requirements, especially for elbow sections. The geometry of all sections is expressed in parametric surface forms in terms of the parametric coordinates u and w, using CAD geometric dimensions. The overall waterslide configuration can be expressed mathematically as X\u00f0u;w\u00de \u00bc XN i Xi ui;wi (8.110) where Xi(ui, wi) is the parametric equation of the ith flume section. N is the total number of sections. Geometric representation of a waterslide in flume sections: (a) assembled configuration and (b) individual flume sections"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001972_ilt-03-2015-0034-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001972_ilt-03-2015-0034-Figure9-1.png",
        "caption": "Figure 9 Pressure evolution for different velocity at 40\u00b0C and 0.08 MPa (EHD calculation)",
        "texts": [
            " This figure clearly shows that there is a creation of a vortex in the axial groove, and the fluid velocity is significant at the lower part of the Figure 5 Mesh bearing 16 nodes 14 nodes 72 nodes 103 nodes 20 nodes Thermal conductivity of the bushing KB (W/m K) 45 Specific heat of the bush CB (J/kg K) 380 Bushing thermal expansion coefficient B (10 5 K 1) 1.2 Bushing Poisson coefficient B 0.33 Bushing Young modulus EB (104 MPa) 12 Turbulent flow behavior in plain journal bearing Nadia Bendaoud, Mehala Kadda and Abdelkader Youcefi D ow nl oa de d by F lo ri da A tla nt ic U ni ve rs ity A t 2 2: 59 0 6 M ar ch 2 01 6 (P T ) bearing and takes significant values for a bearing for working rotation speeds 15,000 rpm, that is to say for the turbulent regime (Figure 8). Figure 9 shows the evolution of the pressure for radial load at 2,000N and for rotational velocity from 6,000 to 15,000 rpm. Pressure distribution of the median plain cylindrical journal bearing for rotational velocity ranging from 6,000 to 15,000 rpm and for radial load 2,000 N is illustrated in Figure 10. The pressure field in two dimensions within the oil film in the axial and angular directions is presented in Figure 11. The bearing is subjected to a load of 2,000 N and operating at a rotational speed of 1,000 rpm (Laminar regime with Re 1,975"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000763_gt2015-43519-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000763_gt2015-43519-Figure8-1.png",
        "caption": "Figure 8: Cross sectional area A0 for zero density (in blue)",
        "texts": [
            " Indeed, interplate distance and interference directly affect the display of the bristle pack, and the behavior of the latter will differ significantly. As a first boundary condition, an orificeflowmeter model will be used to simulate a \u201cseal\u201d with a density of zero bristles per mm. For compressible flows, the mass flow rate is given by [14] : \ud835\udc5a = \ud835\udc36\ud835\udc51\ud835\udc340 2\ud835\udf0c\ud835\udc43\ud835\udc4e\ud835\udc56\ud835\udc5f \ud835\udc58 \ud835\udc58\u22121 (\ud835\udc43\ud835\udc5c\ud835\udc56\ud835\udc59/\ud835\udc43\ud835\udc4e\ud835\udc56\ud835\udc5f )2/\ud835\udc58 \u2212 (\ud835\udc43\ud835\udc5c\ud835\udc56\ud835\udc59/\ud835\udc43\ud835\udc4e\ud835\udc56\ud835\udc5f )(\ud835\udc58\u22121)/\ud835\udc58 (6) - A0 is the cross sectional area of the brush seal without bristles. It is made explicit on Figure 8. 6 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/85152/ on 01/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use - Cd is the discharge coefficient. The definition is the ratio between the mass flow measured at the end of the nozzle to the one of an ideal nozzle. The ISO/ASME empirical formula is [15] : \ud835\udc36\ud835\udc51 = 0,5959 + 0,0315 \ud835\udefd2,1 \u2212 0,184 \ud835\udefd8 + 0,039 \ud835\udefd4 1\u2212 \ud835\udefd4 \u2212 0,0337\ud835\udc51\u22121 \ud835\udefd3 + 91,71 \ud835\udefd2,5 \ud835\udc45\ud835\udc52 0,75 (7) - \u03b2 = d/D, as d is the equivalent diameter of the throat, and in this case study, D is the internal diameter of the air cavity of the bench"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.26-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.26-1.png",
        "caption": "Fig. 35.26 Oil jet lubrication.",
        "texts": [
            " Bearings and applied technology 543 Oil lubrication is generally used where high speeds or operating temperatures prohibit the use of grease, when it is necessary to transfer frictional heat or other applied heat away from the bearing, or when the adjacent machine parts, for example gears, are oil lubricated. Oil bath lubrication is only suitable for slow speeds. The oil is picked up by rotating bearing elements and after circulating through the bearing drains back to the oil bath. When the bearing is stationary, the oil should be at a level slightly below the center of the lowest ball or roller. An application is shown in Fig. 35.25. At high speeds it is important that sufficient oil reaches the bearing to dissipate the heat generated by friction and oil jets provide an effective method (Fig. 35.26). The illustrations in this section (Fig. 35.10e35.26) are reproduced by kind permission of SKF (UK) Limited e www.skf.co.uk. 544 Manual of Engineering Drawing When bearings fail, they can bring equipment to an unscheduled halt. Every hour of down time due to premature bearings failure can result in costly lost production in a capital intensive industry. Substantial investment in research and development has resulted in the manufacture of bearings of the highest quality. Quality alone cannot guarantee trouble-free bearing operation since other factors may affect life span including the following: 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.28-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.28-1.png",
        "caption": "FIGURE 8.28",
        "texts": [
            " The physical position of the spring is shown in Figure 8.26. The spring rate and the damping coefficient were 100 lb/in. and 10 lb/(in./sec), respectively. The free length of the spring was 5.5 in. Note that when the shock was fully extended to its maximum length, the spring length was 4 in., which implies a 150 lb preload. The dynamic simulation (Case A) was carried out assuming a racecar speed of 4.74 mph. The shock travel is shown in Figure 8.27. Note that the shock length was allowed to vary between 7.3 in. and 9 in., as shown in Figure 8.28. This means the shock travel obtained in this dynamic simulation (Case A) is acceptable and the design is safe. Another scenario (Case B) was created where a modified profile cam with a larger hump (4.35 in.) was used, as shown in Figure 8.29. To avoid resonance, a segment velocity (Figure 8.30) was assigned to ensure adequate time for the suspension to return to the equilibrium state between each hump. As can be seen from the resultant shock travel graphs (Figures 8.31(a) and (b)), in Case B the shock was compressed too much and the shock travel exceeded the permitted range, which might have led to a part failure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002253_ecce.2014.6953966-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002253_ecce.2014.6953966-Figure9-1.png",
        "caption": "Fig. 9. Geometry of the second inset machine",
        "texts": [
            " The machine parameters are computed at f=500 Hz freezing the permeability and adopting (8). The real and imaginary parts of the parameters in (4) have been separated as: Some preliminary experimental results are reported in this Section. In particular the high frequency parameters have been measured on a second inset PM machine. The measured values are compared with the result of finite element simulations. The second inset machine is a 4-pole 24-slot prototype available in the laboratory [14]. Fig. 9 shows a sketch of the lamination geometry. The prototype is characterized by unequally distributed magnets in order to reduce the torque ripple [19]. Therefore also the iron teeth between the magnets have different width. Fig. 10 shows a picture of the rotor mounted in the prototype. The PMs are subdivided in four segments in order to facilitate the skewing and to reduce the PM losses. The four segments are clearly visible in Fig. 10. Fig. 11 shows the test bench used to measure the high frequency parameters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003044_cipech.2014.7019080-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003044_cipech.2014.7019080-Figure1-1.png",
        "caption": "Figure 1. Odometry Calculation",
        "texts": [
            " The system can be extended to other configurations. By properly controlling the speed of the wheels, omni wheeled robots can travel in any direction without changing their orientation.(Explained in Section III ). Thus, while discussing about odometry calculations, we need to take 2 angles into consideration \u2013 orientation angle and the travelling angle. As the robot can change its orientation as well as travel in any direction, both these angles are essential to determine to estimate the position of the robot at any instant. As shown in figure 1, we call as the main orientation angle of the robot and as the instantaneous travelling angle of the robot. The is obtained by integrating the angular velocity given by the inertial measurement unit. It can also be made more accurate by combining it with external magnetometer sensor. The 2 rotary encoders fitted perpendicular to each other provide fixed number of pulses per revolution and hence by knowing the diameter of the wheels on which the encoders are fitted, we can find net distance travelled by each encoder"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000984_j.optlastec.2014.11.015-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000984_j.optlastec.2014.11.015-Figure1-1.png",
        "caption": "Fig. 1. Draft of a typical cross section of the welded components: the weld seam joins the nozzle seat to the valve body. Weld bead profile and its geometrical features (WM: martensitic weld width, WF: ferritic weld width, S: resistance length) are magnified on the right.",
        "texts": [
            " This can explained considering that the melt pool shape generated on real components does not change respect to those obtainable with standard specimens (according to ISO 11146:1999). Consequently joining these dissimilar stainless steels without defects is extremely desirable with the generation of welds which can fulfill both technical and economical aspects. The inside diameter of the outer shell and the outside diameter of the seat are machined to \u00d87.50070.025 mm and \u00d87.45870.015 mm respectively to have a clearance fit between them when the shells are assembled. The geometrical features characterizing the weld seam profile are defined in Fig. 1: WM represents the weld width on the martensitic material while WF represents the one on the ferritic stainless steel. The required weld resistance of 3.8 kN is guaranteed by the length of the melt pool at the material interface, here defined as resistance length S, since the joint is supposed to fail under the action of shear stresses. Energy Density (ED) is frequently used, especially in continuouswave laser processes, to incorporate the three main process variables and express the amount of energy flowing per unit area inside the material. ED can be calculated as follows: ED\u00bc P vU\u03c6 \u00f0J=mm2\u00de \u00f01\u00de where P is laser power (in W) describing the intensity of the lasermaterial interaction, v is welding speed (in mm/s) determining the irradiation time and \u00d8 is the laser spot diameter (in mm) defining the area through which the energy flows into the material. In order to generate the required resistance length (as defined in Fig. 1) S4300 mm, empirical models can be adopted to preliminary select the process parameters [15]. In the present case the following combination of parameters was used to generate ED\u00bc40 J/mm2, according to Eq. (1): laser power of 800W, welding speed of 66 mm/s and spot diameter of 0.3 mm. However, experimental practice shows that welding with the selected ED at the interface of the two components with normal laser incidence, induces micro-cracks on the seam, as shown in Fig. 2. Lower ED values do not guarantee the requested resistance length, while higher values increases the number of micro-cracks on the seam"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure21-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure21-1.png",
        "caption": "Fig. 21. Effect of Brake Application",
        "texts": [
            " for 0-3g deceleration was the highest permissible. Communications Mr. G. S. BOWER, B.Sc., Ph.D., A.M.I.Mech.E., wrote that on the assumption that the brake lining behaved elastically, so The forces acting on an element rd8 of the shoe at C would be:that the normal pressure was proportional to its compression by the dnun circle (Watt 1922)*, and that the frictional force on an element of the shoe was proportional to the normal pressure, the virtual radius Zr could be calculated, and the action of the shoe predicted. In Fig. 21, the outline, radius 7, of the shoe and of the drum Radial force acting towards centre 0 = pr cos Bd8. Frictional tangential force acting in direction of rotation of drum, and perpendicular to OC = ppr cos 8dO. Frictional moment about 0 due to latter (positive anti- clockwise) = ppr2 cos Ode. before pressure was applied was represented by the circle tvith centre 0 and radius OB. When the brake was applied, the drum centre moved from 0 to 01, a distance a along the line BO of closest approach. Thus, the greatest compression of the lining was BA = a"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002042_arso.2015.7428209-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002042_arso.2015.7428209-Figure1-1.png",
        "caption": "Fig. 1. The OpenROV with coordinate frames",
        "texts": [
            " We refer to [5] for a detailed formulation of the kinematics of underwater robots. As both the ROV actuators and the camera actuators affect the robot in body coordinates we will use body velocities to describe the robotic motion, i.e., V B 0b = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 u v w p q r \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 . (1) Each of these velocities corresponds to a direction, also defined in the body coordinates of the ROV, given by the vector \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 x y z \u03c6 \u03b8 \u03c8 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 . (2) 978-1-4673-8029-4/15/$31.00 We consider bilateral teleoperation of a ROV that is a OpenROV with an on-board camera, as seen in Figure 1. The operator gives commands through the master haptic manipulator which is connected to a personal computer. We use Phantom Omni haptic device from SensAble Technologies, as seen in Figure 2, which allows for force feedback. The control signals are sent from the PC to the on-board computer through a cable. The setup described above calls for the integration of two rather distinct operation modes: i) accurate control of the on-board camera; and ii) locomotion of the vehicle in a possibly very large workspace"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure4.30-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure4.30-1.png",
        "caption": "Fig. 4.30 Efficiency via the input power for various parameters A. a Cartesian coordinates, b projective coordinates",
        "texts": [
            " a Projective coordinates, b Cartesian coordinates 134 4 Two-Port Circuits Therefore, m1 P1 \u00bc P1 1 \u00f01 P0\u00de P0\u00f01 P0\u00de \u00f01 P0\u00de \u00bc A1: Analogously, for the value A2, we get the load power P2 1 and m2 P1 \u00bc P2 1 \u00f01 P0\u00de P0\u00f01 P0\u00de \u00f01 P0\u00de \u00bc A2: Therefore, the load power change has the view m21 P1 \u00bc m2 P1 m1 P1 \u00bc P2 1 1 P0 1 P1 1 1 P0 1 \u00bc A2 A1 : \u00f04:77\u00de So, there is a strong reason to introduce a specific index in the form P1 1 P0 \u00bc P1 PSC 0 P0 : \u00f04:78\u00de The denominator, as the value PSC 0 P0, shows the load degree of a two-port. Therefore, this index gives more information about a running regime than simply normalized load power. 4.6.3 Influence of Losses on the Efficiency Using (4.73), we get the following efficiency expression KP\u00f0P0\u00de \u00bc P1 P0 \u00bc 1 P0 P0 A \u00f01 P0\u00de2 P0 : \u00f04:79\u00de Let us consider the dependences KP\u00f0P0\u00de for various effectiveness parameters A in Fig. 4.30a. These dependences represent a bunch of hyperbolas passing through the common point PSC 0 \u00bc 1, P1 \u00bc 0 corresponds to SC regime. At OC regime we have KP \u00bc 0; POC 0 \u00bc th2c \u00bc A 1 A : Let us note the hyperbolas with the characteristic effectiveness parameters A. If A \u00bc 1, then the power POC 0 \u00bc 0. In this case, the hyperbola degenerates into two 4.6 Effectiveness Indices of a Two-Port with Variable Losses 135 straight lines, which determine the point S of intersection; the corresponding efficiency KP \u00bc 1. If A = 0, we obtain the hyperbola too. Change of losses Let us consider the projective coordinates in Fig. 4.30b. The bunch of our hyperbolas represents the bunch of ellipses passing through the other common point KP \u00bc 1. The tangential straight lines correspond to this point KP \u00bc 1 and the point PSC 0 \u00bc 1. Therefore, for A \u00bc 1, the hyperbola degenerates into the infinitely remote straight line 1. The above point S is the pole too; the straight line, corresponding to A \u00bc 1, is the polar. In this case, we have the symmetry of the points C1, B1 relatively to the infinitely remote straight line1 or the base points Q; S"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003329_icsai.2016.7810962-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003329_icsai.2016.7810962-Figure1-1.png",
        "caption": "Figure 1. Optimal hyperplane of SVM",
        "texts": [
            " The SVM method is based on structural risk and statistical theory, which can well solve the problems of small samples, high dimensions and non-linearity, but uses limited sample information. In pattern recognition, the SVM is mainly study linearly separable problems. Suppose there is a hyperplane H: w , which can correctly separate the samples. At the same time there are two hyperplanes, both parallel to H, suppose H1 and H2: w \u22c5 x + b =1 w \u22c5 x + b = \u22121 (1) SVM firstly needs to get an optimal hyperplane through training, the optimal hyperplane as shown in the Figure 1. The circles and squares in Figure1 represent two different samples, each of the circles and squares represents a sample point of corresponding space. H is a hyperplane that searched by the algorithm, and its direction is vertical to normal vector. H1 and H2 both parallel to H and cross the sample points hyperplane that is nearest to H. The distance between them known as classified distance. So that the H\u2019s nearest samples just fall on H1 and H2, and the samples are support vectors. In the Figure 1, w represents the normal vector of hyperplane H, b represents segmentation threshold. Then all of the training samples will be located outside of H1 and H2, which means the following constrains are satisfied: w \u22c5 xi + b \u22651 yi =1( ) w \u22c5 xi + b \u2264 \u22121 yi = \u22121( ) 2 The task of the SVM is to find a hyperplane like H that can separates the samples into two parts accurately and make the distance between H1 and H2 is the largest. As shown by Figure 2, fault pattern recognition based on SVM generally has five steps: machinery operation, signature detection, fault feature extraction, SVM fault pattern recognition, maintenance and decision making [6][7]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000685_icit.2015.7125173-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000685_icit.2015.7125173-Figure4-1.png",
        "caption": "Fig. 4 Open circuit flux distributions. (a) conventional, (b) -15 degrees, (c) - 7.5 degrees, (d) 0 degrees, (e) 7.5 degrees, (f) 20.625 degrees.",
        "texts": [
            " This is because for different slot/pole number combination, the flux gap openings to achieve the maximum winding factor are different, e.g. for the 12s/10p, the E-core configuration gives the maximum winding factor while for the rest of the slot/pole number combinations, the modular topologies gives the maximum winding factor. However, the winding factor is not the only quantity affecting the final output torque and the next sections will study the influence of other important parameters. The open-circuit flux distributions are given in Fig. 4 for various values of the flux gap, and compared with the conventional machine. The radial components of the airgap flux density due to PMs are given in Fig. 5 for the conventional, C-core, E-core [with the highest width of the unwound tooth as shown in Fig. 4 (d)] and modular topologies. Only the results for 12s/10p are shown because other slot/pole number combinations give similar results. The radial component of the airgap flux density is non-sinusoidal with high peaks due to double salient nature of the SFPMM. A spatial harmonic analysis of radial component of open circuit airgap field distributions is shown in Fig. 5 b). The conventional machine stands out with respect to the modular ones by having only even order harmonics. The removal of half of the magnets leads, in case of the modular machines, to a complete change in the harmonic spectrum"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002715_ijmmp.2016.078055-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002715_ijmmp.2016.078055-Figure1-1.png",
        "caption": "Figure 1 (a) Schematic view of laser processing setup utilised for experimentation and (b) diode laser surface melting setup (closer view) includes laser beam of 4 kW power impending from focusing lens. Laser beam shroud with N2 gas coming from shroud",
        "texts": [
            " Comprehensive characterisation of microstructure of melted layer was carried out and correlated to enhancement in corrosion resistance as compared to that of untreated counterpart. The material used in the present work was 7075-T651 aluminium alloy heat treated to peak-aged condition. Table 1 shows the nominal chemical composition of the alloy. A 6.35 mm thick plate with polished surface (utilising fine-800 grit SiC paper) was irradiated by laser source. The 7075-T651 aluminium alloy plate was irradiated under a multi-mode (Gaussian in fast axis and top-hat in slow axis) diode laser beam of size 20 \u00d7 5 mm. Figure 1 shows the laser processing setup utilised for experimentation. The setup constitutes a 6-kW maximum power deliverable continuous wave fibre-coupled diode laser system integrated to 8-axis robotic workstation. It is pertinent to mention here that diode laser whose wavelength is 890\u2013980 nm can potentially induce high coupling efficiency to aluminium-based materials and thereby facilitate high process efficiency (TRUMPF, 2010). The sample was fixed to the working table and the laser beam carried through fibre connected to the sixth arm of robot is traversed at a constant laser power of 4 kW with scanning speed comprising of 10 mm/s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000075_gt2012-68354-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000075_gt2012-68354-Figure1-1.png",
        "caption": "Figure 1. The principle of the brush seal",
        "texts": [
            " Even though the rotor was accelerated to high speeds up to 19500rpm, the produced temperature overshoots in the seal/rotor contact zone have caused no deterioration in either the materials or the oil. This work is part of the European Union funded research programme ELUBSYS (Engine LUBrication System TechnologieS) within the 7th EU Frame Programme for Aeronautics and Transport (AAT.2008.4.2.3). Brush seals [2], [4] are sealing mechanisms which consist of a very large number of bristles that are wound around a rod (core wire). The rod is circular and the bristle package is retained by the backing support plate and the front plate on both sides (Figure 1). A clamping tube holds the bristle package tight around the rod. The seal is static and the bristles package may run in contact or even with an overlap to the rotating part. Air or other gases may be used for pressurizing the seal thus designated as gas/gas or as gas/liquid seal. The bristle package may be composed by metal, Kevlar or other fiber materials. MTU Aero Engines is one of the brand names worldwide in manufacturing brush seals of metallic or Kevlar types. Depending upon the seal type, the seal density may be 1 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure11.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure11.9-1.png",
        "caption": "Fig. 11.9 a Power supply system with resistance R1. b Its geometric model",
        "texts": [
            " 11.1 Analysis of Load Voltage Stabilization Regimes 323 11.2 Given Voltage for the First Variable Load and Voltage Regulation of the Second Given Load We consider again the circuit shown in Fig. 11.1. Let the first load voltage V1\u00bc be stabilized. But for all that, the first load resistance may be both positive R1 [ 0 and negative R1\\0. Also, the second constant load resistance is positive R2 [ 0. For example, the circuit in Fig. 11.8 corresponds to the positive load R1 [ 0 and PWM regulators in Fig. 11.9a conform to the negative load R1\\0. We rewrite Eq. (11.1) Ri R1 \u00f0V1\u00de2 \u00fe Ri R2 \u00f0V2\u00de2 \u00fe V V0 2 2 \u00bc \u00f0V0\u00de2 4 ; which correspond to a surface with a parameter R1 in the coordinates V1, V2, V3. If R1 [ 0, this expression represents a sphere (ellipsoid) similarly to the circle in Fig. 11.2. The both loads consume energy; the voltage source V0 gives energy. If R1\\0, we get a one-sheeted hyperboloid in Fig. 11.9b [3]. The first load, as a constant voltage source V1\u00bc, gives energy. In addition, the voltage source V0, as energy storage, may consume and give back energy. The corresponding direction of the current I0 determines these regimes. For different values R1, our expression represents a bunch of spheres or hyperboloids. If R1 \u00bc 1, as the open circuit regime, the corresponding surface degenerates into a cylinder. 324 11 Stabilization of Load Voltages Let us now return to given operating regime; that is, V1 \u00bc V1\u00bc; R2 \u00bc const: For realization of this regime, it is necessary to change the transformation ratios n1; n2 in some coordination"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000039_ijaac.2017.083296-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000039_ijaac.2017.083296-Figure1-1.png",
        "caption": "Figure 1 Schematic figure of twin rotor with main and tail propulsion",
        "texts": [
            " The rotor system can be placed in a helicopter horizontally, just as main rotors which make it possible for helicopters to take off in a vertical direction. Also, this system can be implemented vertically, just as tail rotors which make it possible for helicopters to take off in a horizontal direction by creating appropriate force to compensate the torque effect. Twin rotor is controlled by two inputs, namely u1 and u2. Mutual coupling is one of the main features of twin rotor. The positions of helicopter bars are measured by incremental encoders which provide the relative positions from signal. According to Figure 1, twin rotor consists of two thrusts, i.e., primary and secondary, which are run by two independent primary and secondary DC motors. These thrusts are perpendicular to each other and are attached to each other by a bar which can rotate freely on horizontal and vertical planes. Pitch and yaw angles can be justified by changing the voltage of primary and secondary motors, respectively; this is done by controlling the rotation speed of primary and secondary thrusts. Designing an effective controller to track pitch and yaw desirable angles is a difficult task due to the coupling between thrusts"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001207_gt2014-25290-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001207_gt2014-25290-Figure5-1.png",
        "caption": "Figure 5: Rotating out-of-unbalance forces is the 2:unb \u00d7 matrix whose columns are the eigenvectors evaluated at the locations and directions of the entries in :",
        "texts": [
            " \u2022 Omission of from equation (2). The vector comprises the unbalance forces, which are defined as follows, = +,-. ,/. ,-0 ,/0 \u22ef\u22ef ,-2unb ,/2unb6 (4) where the unbalance forces ,-7, ,/7 8 = 1, 2, 3, . . . , :unb are assumed to be concentrated at disk locations 8 =1, 2, 3, . . . , :unb: ,-7 = ,;<= sin < + @; (5a) ,/7 = \u2212,;<= cos < + @; (5b) where ,; is the unbalance (in kgm) at disk no. 8 and @; is its angular position with respect to the arbitrary reference marker that makes angle < with the downward vertical (Figure 5). = CDD GI7 \u22ef GI7 \u22ee \u22ee \u22ee JKK KL (6) 3 Copyright \u00a9 2014 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use is the vector of the SFD forces on J in the x, y directions which are nonlinear functions of the displacement and velocity of J relative to B. The SFD bearing considered for this illustrative study was single land and end fed with oil of viscosity 0.0049 Nsm -2 at a pressure of 3 bar (gauge). The bearing diameter and radial clearance were 220mm, 0",
            " In this research three balancing planes are used and so: df = CD DD DD E,n cos @o ,n sin@o ,n= cos @o=,n= sin@o=,np cos @op,np sin@op JK KK KK L de = CD DD DD E ,n sin@o \u2212,n cos @o ,n= sin@o=\u2212,n= cos @o=,np sin@op\u2212,np cos @opJK KK KK L (15a, b) Notice that the actual unbalances , , ,=, ,p,\u2026 will generally reside in more than three planes and the three equivalent unbalances ,n , ,n=, ,np will not be at the locations of , , ,=, ,p. However, the angular displacement @o; is with respect to the same arbitrary reference marker used in Figure 5 (the line that makes angle < with the downward vertical). It 4 Copyright \u00a9 2014 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use is noted that the locations of ,n , ,n=, ,np are the balancing planes, where unbalances equal and opposite to ,n , ,n=, ,np will be applied once these are estimated. From equation (10): N OR = \u2211 qOR\\ ] ]\u0302_ cos `< + qORa ] sin `< (16) where qORr ] = `\u03d6 NORt ] , qORa ] = \u2212`\u03d6 NORr ] (17a, b) Hence, equations (12a,b) can be expressed as: qOR\\ = POR < u +POR d < <= df , V = 1\u2026:W (18a) qORa = \u2212v POR < w +POR d < <= de x, V = 1\u2026:W (18b) where POR < , POR d < are the mobility matrices: POR = < cOR < , POR d = < cOR d < (19a,b) From equations (18a,b), the synchronous components of the SFD forces can be expressed in terms of the unbalance and the synchronous components of the velocity measurements at one of the sensor locations",
            " The first technique is Least Squares Fit (LSF) [10]: y = y y j|y (33) yf = yW yW j|yW (34) The second technique uses Singular Value Decomposition (SVD) of y and yW [10], which results in the following pseudo-inverse: y = , yf = W W W (35, 36) where, if y and yW are of size \u00d7 :: and W are \u00d7 matrices whose columns are the eigenvectors of y y and yW yW respectively; and W are : \u00d7 : matrices whose columns are the eigenvectors of y y and yW yW respectively; is the : \u00d7 matrix given by: = diag j \u22ef j !! ! (37, 38) 5 Copyright \u00a9 2014 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use where ,\u2026 are the K non-zero singular values of y ; W is defined in a similar fashion. The estimated equivalent unbalance ,n;, 8 = 1,2,3 and its angular displacement @o; relative to the arbitrary reference marker that makes angle < with the downward vertical (Figure 5) can be estimated from either dW (equation (27)) or de (equation (28)): ,n; = ,n; % ` @o; = + ,n; `V: @o; = (39) @o; = atan2 +Gn.7 ]U n7Gn7 , Gn7 \u00a0] n7Gn7 6 (40) Hence, the balancing force vector \u00a1 can be structured as follows: DE <=,n `V: < + @o + \u00a2 \u2212<=,n % ` < + @o + \u00a2 <=,n= `V: < + @o= + \u00a2 \u2212<=,n= % ` < + @o= + \u00a2 <=,np `V: < + @op + \u00a2 \u2212<=,np % ` < + @op + \u00a2 JK KL The numerical validation of the proposed method involves the inclusion of an additional unbalance term \u00a1 \u00a1 to the right hand side of equation (1) and the solution of the resulting equation for the response, as described in section 3, where \u00a1 is the 6 \u00d7 matrix whose columns are the eigenvectors evaluated in the x, y directions at the balancing locations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure1.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure1.19-1.png",
        "caption": "Fig. 1.19 Buck converter",
        "texts": [
            " It is determined by how to set the value changes; that is, by increments, ratio and so on. The situation becomes complicated even more, if such a power supply system contains two loads with individual voltage regulators shown by dash lines in Fig. 1.18a; the interference of these loads takes place. VR2 R 1 V 1 VR1 VR i n 2 n 1 V 0 R 2 V 2 0 1 2 3 4 5 0.00 0.25 0.50 V 0 V 1 n 1 (a) (b) Fig. 1.18 a Power supply system with limited capacity voltage sources. b Example of its regulation characteristic 22 1 Introduction 1.6.2 Buck Converter Let us consider a buck converter in Fig. 1.19. The expression of the static regulation characteristic for the continuous current mode of a choke L with a loss resistance R has the known view [5, 10] VL \u00bc V0 1\u00fe R RL D \u00bc V0 1\u00fe\u00f0r\u00de2 D; \u00f01:37\u00de where D is the relative pulse width and \u00f0r\u00de2 is the relative loss. Let us express (1.37) in the relative form. There are some variants. First, it is possible to introduce the value VL \u00bc VL=V0. Then VL \u00bc 1 1\u00fe\u00f0r\u00de2 D: \u00f01:38\u00de This expression is not a pure relative because it contains the value \u00f0r\u00de2. Let D be changed, D1 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003759_gt2016-56951-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003759_gt2016-56951-Figure6-1.png",
        "caption": "FIGURE 6. ROTOR INSTRUMENTATION - MEASURING POSITIONS",
        "texts": [
            " The thermocouples, which are led through the covers, are welded steamtight and subsequently laid to the sensing locations. Figure 5 shows an example of the superficial distribution of thermocouple wires at the drive end. The individual thermocouples are led through the above-mentioned cover plate. The wires are superficially fixed by spot-welded straps. The thermocouples are led into eroded holes with different depths in different radial and circumferential positions to the appropriate measuring positions. Figure 6 shows the resulting distribution on both rig sides, projected in one plane. In total, 26 thermocouples with an outer diameter of 1 mm are implemented. Figure 6(a) shows the instrumentation of 12 thermocouples underneath an investigated tandem seal design as already in- troduced in [5], whereas Figure 6(b) shows the instrumentation of 14 thermocouples underneath a schematic three-seal arrangement. The four different radial measurements planes are located in a range of 2 up to 9.5 mm underneath the rotor surface. The whole development process was supported by FEM calculations in order to guarantee a safe test operation. Despite the 5 Copyright \u00a9 2016 by Siemens Energy, Inc. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 09/03/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use high thermal and centrifugal loads, material stresses and stability are uncritical for the described rotor design"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000113_ijnm.2017.082406-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000113_ijnm.2017.082406-Figure3-1.png",
        "caption": "Figure 3 The definition of cutting forces\u2019 directions",
        "texts": [
            "4538 Image about chip morphology with the mount of micro-milling cutter overhanging L = 12 mm; spindle speed n = 39,680 r/min; axial depth of cut ap = 15 \u03bcm; feed per tooth fz = 3.0 \u03bcm/z is shown in Figure 2 and long strip chip are obtained. The processing length of groove is 30 mm with dry cutting and air-cooler during experimental processing. Kistler 9256CQ1 is used to measure the cutting forces. Fx, Fy and Fz are cutting forces of different directions. In the micro-milling grooving orthogonal experiment, X direction is the cutting feed direction, and Y direction with perpendicular to the feed direction is the main cutting force direction, as shown in Figure 3. The experiment results (see Table 2) indicate that the main cutting forces are larger than feed forces within the scope of the same experimental cutting parameters. It can be seen from Figure 4, when the mount of micro-milling cutter overhanging varies between 12 mm and 16 mm, cutting forces are little influenced by the mount of micro-milling cutter overhanging; as the extended length of micro-milling cutter increases, cutting forces first increase and then decrease, the variation trend of which is quite slow"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000017_022050-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000017_022050-Figure1-1.png",
        "caption": "Figure 1. The technological scheme of the sod seeder of strip sowing (a) and its coulter group (b): 1\u2013packing roller, 2 - seed coulter, 3 - fertilizer coulter, 4 - hopper for seeds and fertilizers, 5 - disc milling cutter; 6 \u2013 suspension mechanism of coulter group; 7 \u2013 protective casing.",
        "texts": [
            " The aim of the study is to theoretically determine the parameters of the opener group of a seeder for strip sowing of grass seeds into sod with the introduction of a starting dose of mineral fertilizers. To increase the efficiency of the process of direct strip sowing of grass seeds into the sod with SDK seeders, it was proposed to remove the working bodies for seeding seeds and embedding fertilizers from the operating zone of disc cutters. In this case, it is necessary to combine the fertilizer and seed coulters into a single coulter group. Removing the coulter group from under the protective casing of the milling furrow-opener of the seeder (figure 1, b) will reduce the soil sticking to the space under the casing and exclude the influence of casing vibrations on the coulters from impacts of clods earth. Reducing sticking of the disc cutter casing is important in conditions of high humidity, and eliminating the influence of casing vibrations on the coulters from impacts of clods of earth will improve the uniformity of the seeding depth of mineral fertilizer granules and seeds. This method includes pre-sowing strip processing of the sod with disc cutters, leveling the soil, applying mineral fertilizers and sowing seeds on a compacted seeding bed along the axis of the treated strip 0.01-0.02 m above the depth of fertilization (figure 1, a). In this case, a double opener is installed for fertilizers and grass seeds. Torsion springs are used to mount the coulter group. When driving in a milled strip of soil, the fertilizer coulter creates a furrow into which mineral fertilizers are sown. After the coulter has passed, the furrow walls crumble and cover the fertilizer. This ensures an optimal soil layer between mineral fertilizers and seeds. Further, the opener compacts the seedbed over the granules of mineral fertilizers and grass seeds are sown"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002118_1754337116638970-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002118_1754337116638970-Figure4-1.png",
        "caption": "Figure 4. (a) Pendulum arrangement to measure the \u2018pick-up weight\u2019 and (b) and bat suspension arrangement to measure the \u2018polar moment of inertia\u2019.",
        "texts": [
            " The low-frequency residuals are a direct measure of rigid body mass properties of the bat due to the freely suspended boundary conditions. The scaled mode shapes were used to calculate the energy absorbed at each excitation point on the blade.7 The MoI about two axes was determined and used to indicate the differences in rotation characteristics of the bat designs. By turning the bat into a physical pendulum11 and using the set-up described by Eftaxiopoulou et al.,15 the MoI about the AA axis of the bat, called the \u2018MoI pick-up weight\u2019 from now on, was calculated as shown in Figure 4(a). The \u2018MoI pick-up weight\u2019 along with the bat\u2019s mass determines the amount of effort required by the batsman to swing and control the bat during the swing.11 In addition, the \u2018polar MoI\u2019 was defined as described by Brody11 calculating the resistance to rotation of the bat about the YY axis (Figure 4(b)). The \u2018polar MoI\u2019 is a measure of the energy imparted at the edges of the bat and directly related to the ability of a miss-hit at the edge to travel further. Table 1 shows a list of all the bats tested along with sample design parameters. As can be seen, the newer bats (manufactured after 2009) are thicker, heavier and have their mass distributed closer to the toe when compared to the earlier designs. The results related to the first four distinct mode shapes and frequencies for each bat are summarised in Table 2 showing that the first four modes of vibration for all bats occur at similar frequencies with the exception of Bat 1 whose modes of vibration occur at a noticeably lower frequency than those of the other bats"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001128_robio.2014.7090523-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001128_robio.2014.7090523-Figure2-1.png",
        "caption": "Fig. 2 The kinematic model of the sys",
        "texts": [],
        "surrounding_texts": [
            "A. Kinematic model of wheelchair-manip The mobile manipulator that used for doo shown in Fig. 1. The manipulator is mounted on the r wheelchair. And the non-holonomic prob taken into consideration in the system [15]. The speed of the end-effector that is movement of the wheelchair can be express mme qJp = where [ ]Te zyxp \u03b3\u03b2\u03b1= is th end-effector, and \u2212 + == v q r r m m m ( ( \u03b8\u03b8 \u03b8\u03b8 \u03c9 wheelchair\u2019s velocity, in which mv is velocity that parallel to the wheelchair and \u03c9 velocity for the horizontally turning the wh radius of the driving-wheel and L is the dist two driving-wheels. Then l\u03b8 and r\u03b8 are joi of driving-wheels. 26\u00d7\u2208 RJm is Jacobin m velocity of the wheelchair to the velocity of of the manipulator: \u2212 = 10 00 00 00 )cos(0 )sin(1 \u03c6 \u03c6 L L J m 22 )()( yyxx lPlPL +++= where xP and yP are the x-y coordinates of base in b , \u03c6 is the angle of the b , whic ulator system r-opening task is ystem ight side of the lem needs to be induced by the ed as following: (2) e velocity of the LR R l l /) 2/) is the the longitudinal m is the angular eelchair; R is the ance between the nt motor velocity atrix that map the the end-effector (3) (4) the end-effector h is the same as the angle of the wheelchair base on t The forward kinematics of the 6 respect to b can be expressed as f (aa qfp = where ap is the position and orienta end-effector that are expressed in t [ ]Taq 654321 \u03b8\u03b8\u03b8\u03b8\u03b8\u03b8= of the manipulator. Combine the wh together, the forward kinematics of t the world coordinate w can be exp ,( me qhp = m Tp += the variable T cccm yxp ][ \u03b8= orientation of the wheelchair with re transformation matrix m bT transfers frame b to the frame of the wheelc Differentiating both sides of (6) w m m a m e V q hV q hp = \u2202 \u2202+ \u2202 \u2202= [ ]= m a ma V V JJ or p where q [ 54321 \u03b8\u03b8\u03b8\u03b8\u03b8= joint variable of the wheelchair-man By modeling the system as (7), t explicitly entail the admissible spee respect to its non-holonomic constra B. Kinematic model of door open The situation is considered that the the door handle and has achieved a of the door. What\u2019s more the door is the hinge and not opening by movin right. And the door handle is gras there is no relative translation an manipulator\u2019s end-effector and the fixed grasp), and the door handle m door hinge is rotating so the door considered as a planar control prob described above is the most commo door in the domestic environment. Now, we consider the kinemati opening problem [2]. As shown in F the end-effector frame and constra The constrained frame is attached a And the radial direction vector r i position of the frame e and c , he world frame. -DOF manipulator with ollow: )a (5) tion of the manipulator\u2019s he base frame b , and denote joint variables eelchair and manipulator he system with respect to ressed as follow: )aq )( aa c b qf (6) is the position and spect to w , and the the manipulator\u2019s base hair. ith respect to time gives: mmaa VJVJ + qJe = (7) T mmv ]6 \u03c9\u03b8 is the ipulator system. he kinematic model can d of the wheelchair with in. ing manipulator has located fixed grasp of the handle opened by the rotation of g door plane from left to ped by the manipulator, d rotation between the door handle (so called oves in a plane when the opening problem can be lem. While, the situation n case that opening the c constrain of the door ig. 2, e and c denote ined frame respectively. t the hinge of the door. s defined as the relative r can be expressed as bellow: ec ppr \u2212= Let )(\u22c5o denote the orthogonal operator, in to say rro \u22a5)( . Notice that the mo end-effector of the manipulator is constrai direction, so the velocity along the radial di zero, and the following formula can be obta 0=e wT pr Tr is the transpose of the radial direction v the velocity of the end-effector in the world III. CONTROL SYSTEM DESIG In order to open the door successfully, the the control system is shown in Fig. 3. Th mainly contains three controllers, which are controller, orientation controller and the co controller that generates motion for the joi redundant system. Where c dp is the objective velocity for open is expressed in constrain frame c . The te the objective force and torque at the respectively. Transformation matrix c eT an (8) this case, that is vement of the ned at the )(ro rection of door is ined [4]: (9) ector r , e w p is frame. tem N block diagram of e control system the translational ordinated motion nt variable of the eme ing the door that rm c df and c d\u03c4 is non-driving-axis d c mT transfer the manipulator\u2019s end-effector frame an to the constrained frame c . The system uses the 6-axis force force and torque of the end-effector use the velocity controller and o generate the suitable velocity of the coordinated motion controller is corresponding joint velocity of the m of the wheelchair\u2019s two wheels. The scheme is implemented in the constr force/torque is measured in the end- A. Translational Controller The translational controller can co end-effector motion be close to the uses an active compliance con connecting the force of the end-effec spring-mass model [7]. The dynamic hybrid compliance control [5] are used for the end-effec generation. Dynamic hybrid compli firstly estimate the constrained fram constrained frame into driving-axis a driving-axis is defined as this: when the other constrained mechanisms end-effector is constrained by the m movement direction of the manipul same as the constrained motio driving-axis is the axis that tangent manipulator\u2019s end-effector, along w opened. And the non-driving-axis driving-axis that shares the door-open-trajectory plane. The applies a constant driving-velocity a applies a compliance force co non-driving-axis. At the non-dri between force error and position err When opening the door, the geometr then the movement of the end-effe grows unexpectedly. The translati desired velocity c dp along the drivi impedance control along the non-driv to track a desired force c df . Considering the constrained fram the relation between position erro force error c d cc fff \u2212=\u0394 at the described in the following equation: c v c v c pDpMf \u0394+\u0394=\u0394 where vM , vD and vK den damping and stiffness matrix relate t respectively. At the non-driving-axis d mobile platform frame /torque sensor to get the of the manipulator. Then rientation controller to end-effector, after that, a used to compute the anipulator and the speed force and torque control ained frame c and the effector frame e . nstrain the manipulator\u2019s door handle\u2019s motion. It trol method, and can tor with its position by a control and impedance tor translational velocity ance control scheme can e and then decouples the nd non-driving-axis. The opening door or operates , the movement of the echanism, and the actual ator\u2019s end-effector is the n direction, then the to the trajectory of the hich the door could be is perpendicular to the same plane as the translational controller t the driving-axis, and it ntrol method in the ving-axis, the relation or is mainly considered. ic of the door is unknown, ctor will cause the force onal controller track a ng-axis and it apply the ing-axis, the objective is e c , according to [5], r c d cc ppp \u2212=\u0394 and non-driving-axis can be c v pK \u0394+ (10) ote the virtual intra, o the position , the input force error cf\u0394 can obtained from the 6-axis force/torque sensor, and then position error cp\u0394 can be obtained through (10). The control objective is to reduce the position error cp\u0394 so that the contact force will decline. By calculating the equation (10) the cp\u0394 can be expressed as follow [5]: )()()( kpkpkp cc d c c \u0394+=\u0394 (11) At the k-th period, c cp\u0394 is the actual translational velocity expressed in c . So, the desired velocity of the next period can be defined as follow [5]: )()1( kpkp c c c d =+ (12) Differentiating equation (11) and combine with (12), the following formula can be obtained: )()()1( kpkpkp cc c c c \u0394+=+\u0394 (13) Then, at the non-driving-axis, the control law is described as (13). A constant velocity is applied at the driving-axis, so the translational velocity of the manipulator\u2019s end-effector e\u03c5 can be obtained. B. Orientation Controller Considering the orientation controller, an angle impedance control algorithm is applied to control orientation of the manipulator\u2019s end-effector. At the constrained frame c , the relation between the torque c\u03c4 and the angular displacement \u03b8\u0394 can be described as follow [7]: \u03b8\u03b8\u03b8\u03c4 \u0394+\u0394+\u0394= www c KDM (14) where \u03c9\u03b8 =\u0394 is the angular velocity of the manipulator\u2019s end-effector. And wM , wD , wK denote the virtual intra, damping and stiffness matrix relate to the end-effector\u2019s orientation respectively. And c\u03c4 is the torque expressed in c . Based on (14) and replace \u03b8\u0394 , the following formula can be obtained: \u03c4\u03c4\u03c9\u03c9\u03c9\u03c4 dKDM t www c ++= 0 )( (15) So, the velocity of the end-effector in task space can be obtained, which includes the translational velocity and the orientation velocity. C. Coordinated Motion Controller As the velocity of the manipulator\u2019s end-effector is obtained, now it should mapped to velocity of the manipulator and the longitudinal/angular velocity of the wheelchair respectively to control the system to open door. While, the wheelchair mount with a 6-DOF manipulator is a highly redundant system when only the velocity of the end-effector is specified, so, a coordinated motion controller is needed to coordinate the movement of wheelchair and manipulator. Then, combining the kinematic models of the wheelchair and the manipulator together by means of Jacobean augmentation, the formula of the most commonly used kinematic control methods with redundancy resolution can be obtain as following [16]: \u03be)( *JJIrJq \u2212+= \u2217 (16) Where [ ]TAC qqq = is the joint variables of the system and \u2217J is the pseudo inverse of Jacobean matrix J , [ ]AC JJJ = is the combination of the Jacobean of wheelchair and manipulator, rJ \u2217 , the first item of (16), denotes a minimum Euclidean-norm solution to the equation qJr = ; \u03be)( *JJI \u2212 , the second item of (16), denotes the homogeneous solution that is in the null space of J , and corresponding to the manipulator\u2019s self-motion without affecting the end-effector motion; the second item is commonly used to satisfy some secondary tasks. It could be to avoid the joint limit or, more commonly, to optimize the manipulability of the robot. So, why we choose to improve the manipulability when opening the door? Firstly, the manipulator is a high stiffness system [17], even small position change at the end-effector will cause big force at the wrist of the robot arm (fixed on the door handler). What\u2019s more the mass of the wheelchair is very big and abrupt change of its velocity will be hard to response for the system. So, the movement of the manipulator\u2019s end-effector and the joint variables should be as smooth as possible. Secondly, the kinematic singularities cannot be avoided in the simple pseudo inverse approach [18]. However, the singularity situation could occur with higher possibility in the redundant system, so when opening the door, we need to improve the manipulability of the system so that the mobile manipulator will not reach the singularity. Thirdly, the pseudo inverse is a minimum Euclidean-norm solution, it will try to make the joint variable to change as small as possible, however, as with our humans, we may want to keep a comfortable position when opening the door, so that we not only use our hand to pull the door\u2019s handle, but also we will adjust our body to make the door opening behavior as comfortable as possible. Because of the above mentioned reasons, the criterion is chosen to optimize the manipulability of the robot when opening the door. The manipulability measure is defined as bellow [18]: )det( TJJw = (17) So, the item \u03be in (16) can be derived from (17), the i-th element in \u03be can be given mathematically as follows [19]: i T i i q JJ q w \u2202 \u2202 = \u2202 \u2202= )det(\u03be (18) )))(()(()det( 1 T i T i TT q JJJ q JJJtrJJ \u2202 \u2202+ \u2202 \u2202= \u2212 , in the above equation )(\u22c5tr denotes the trace of a matrix. Finally, according to [19] the equation to control the system can be obtained as follows: \u03be)( *JJIkpJq e \u2212+= \u2217 (19) where k is a scalar to weight the second term of (19), so that we can control the self-motion of the system when an end-effector path is given. The above method is the most commonly used way to maximize the manipulability when the movement of the end-effector is specified or under constrain."
        ]
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.1-1.png",
        "caption": "Fig. 6.1 (a) Basic shape rolling operation. (b) Force and geometry of fully engaged rolling",
        "texts": [
            " Deformation processes are of utmost importance in today\u2019s industrial mass production operations as these processes enable efficient production and do not rely on lengthy metallurgical processes. In rolling, a metal slab passes through a pair of constantly rotating rolls and is permanently deformed in the roll zone where the rolls compress the metal slab. The friction generated due to the spinning of the rollers moves the metal slab forward. The body thickness is reduced and equalized by the compressive forces during the rolling process (Fig. 6.1a), while the volume is largely maintained. As shown in the half block model (Fig. 6.1b), the metal block will be drawn in between the rolls with a normal force Fn and a friction force Ff with the relation Ff \u00bc \u03bcfFn, where \u03bc is the friction coefficient, and hence, tan(\u03b1) \u00bc \u03bcf. Furthermore, cos(\u03b1) \u00bc 1 \u2013 TR/(2R), where R is the radius of the roller and TR is the thickness reduction of the block after rolling. Along with this, sin(\u03b1) \u00bc [1 \u2013 cos(\u03b1)]1/2 (TR/R) 1/2. Hence, the maximum feasible TR is \u03bcf 2R. In addition, the rolling force (FR) can be estimated by the forming pressure PF length of contact LC and width of the workpieceW: FR\u00bc PFLCW"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002490_6.2016-3199-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002490_6.2016-3199-Figure1-1.png",
        "caption": "Figure 1. A pitch oscillation motion defined as \u03b8 = Asin (\u03c9t) .",
        "texts": [
            " In this work, the stability derivatives of a generic missile are calculated by (1) imposing a forced sinusoidal motion around the center of gravity of the vehicle and (2) from the vehicle indicial responses to a unit step change in the angle of attack and normalized pitch rate. These methods are briefly described in this section. Forced oscillation motions are the most common methods to estimate dynamic derivatives in the wind tunnel and using CFD.12 These estimations, however, depend on the amplitude and frequency of the motion. A pitching sinusoidal motion, as shown in Fig. 1, is defined as: \u03b1 = \u03b10 +Asin (\u03c9t) (1) where \u03b10 and A are the mean angle and amplitude of oscillation, respectively. The value \u03c9 = 2\u03c0f is the angular velocity about the center of gravity. The pitch angle, \u03b8, is equal to the angle of attack for the form of motion considered. The pitch rate is therefore written as: q\u0304 = \u03c9Acos (\u03c9t) (2) The normalized pitch rate is then defined as: q = q\u0304D 2V (3) where D denotes the missile body diameter and V is the free-stream velocity. Notice that pitch rate, q, and angle of attack time rate, \u03b1\u0307, have different effects on the aerodynamic loads"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002864_0954405416661003-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002864_0954405416661003-Figure2-1.png",
        "caption": "Figure 2. Spiroid gear generated by a virtual pinion cutter.",
        "texts": [
            " In order to solve the problems above, modification of the tooth surfaces for helicon gears usually had been performed by the application of ease-off modification, which enables the gear pair to be point contact instead of instantaneous line contact. The design of the raised novel geometry of the spiroid gear is based on the following ideas: 1. The basic geometry of the spiroid gear is imaginary generated by the use of a virtual pinion cutter based on the method of the conventional gear generated by a hob (Figure 2). 2. The novel tooth geometry with skew doublecrowned gear is achieved by the application of ease-off modification methodology of contact path and contact line direction. 3. The modified tooth geometry of gear will be offered as die cavity surface that is manufactured by electrical discharge or CNC machine directly. 4. The geometry of double-crowned gear will be manufactured by precision casting process using such modified geometry of die cavity. Geometry of the imaginary generating fully conjugated spiroid gear In theory, the fully conjugated tooth surfaces of spiroid gear can be obtained by using a spiroid cylindrical pinion as a virtual hob to produce the gear as shown in Figure 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000700_978-3-319-11271-8_17-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000700_978-3-319-11271-8_17-Figure6-1.png",
        "caption": "Fig. 6. Conception of robots formation control with errors \u03b4(j)L{k} and \u03c8(j) d{k}",
        "texts": [
            " Positions of characteristic points of the virtual structure A(j) d , are traced by the WMRs pointsA(j) in the way, that the j-th WMR\u2019s pointA(j)(x(j)A{k}, y(j)A{k}) is going to achieve in the next iteration step the desired position A(j) d (x(j)Ad{k}, y(j)Ad{k}) computed on the basis of the virtual structure position and orientation. Determined trajectories guarantee minimisation of errors \u03b4(j)L{k} and \u03c8(j) d{k}, what results in the trajectories, in which the pointA(j) of the j-th WMR traces the pointA(j) d of the virtual structure. The idea of formation control is shown in Fig. 6. The WMRF control signals were assumed in the form u (j) Fv{k} = kF1\u03b4 (j) L{k} cos ( \u03c8 (j) d{k} ) , u (j) F\u03b2\u0307{k} = kF1 sin ( \u03c8 (j) d{k} ) cos ( \u03c8 (j) d{k} ) + kF2\u03c8 (j) d{k} , (13) where kF1, kF2 \u2013 positive constants. The presented formation control system was discussed in detail in [2,7]. On the basis of the WMRF control signals u(j)Fv{k} and u(j) F\u03b2\u0307{k} were computed an- gular velocities of j-th WMR proper wheels according to equation [ z (j) d2[1]{k} z (j) d2[2]{k} ] = 1 r [ v\u2217M \u03b2\u0307\u2217l1 v\u2217M \u2212\u03b2\u0307\u2217l1 ] [ u (j) Fv{k} u (j) F\u03b2\u0307{k} ] , (14) where v\u2217M \u2013 a maximal define velocity of the pointM , \u03b2\u0307\u2217 \u2013 a maximal defined angular velocity of the WMR\u2019s frame self-turn"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003477_pedes.2016.7914438-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003477_pedes.2016.7914438-Figure5-1.png",
        "caption": "Fig. 5. Field lines for S Q 12= (left) and S Q 60= (right), both for a number of pole pairs p 1= .",
        "texts": [
            " 4 shows the inverse-square dependency of the main inductance on the number of pole pairs. Both Figures (3 and 4) illustrate an almost perfect match between FEM and analytical calculation. This is due to the fact that the FEM model is built under the same assumptions as were used during the analytical modelling process. inductance as a function of number of pole pairs, for S Q 60= . B. Calculation with Consideration of Stator Slots Calculation including the stator slots is performed for the same variables like in Section A. In addition, the stator slotting is considered, see Fig. 5 for exemplary FEM calculations. To introduce a realistic slot opening for different numbers of stator slots, the stator slot opening is defined as follows: so S 60 b 2mm Q = \u22c5 (33) To account for the slot opening effect onto the main inductance, the well-known Carter factor is used (see e.g. [6]) to virtually increase the air-gap-width ( S \u03c4 according to Eq. (1)): C 2 S so so C S K b b K , 5 \u2032\u03b4 = \u03b4 \u22c5 \u03c4 = \u03b1 = + \u03b4 \u03b4\u03c4 \u2212 \u03b4 \u22c5 \u03b1             (34) Using the parameters defined above, the Carter factor varies between C K 1.2558= for S Q 12= and C K 1.1276= for S Q 60= . Fig. 6 shows the comparison of analytical and FEM calculations of the stator main inductance as a function of the number of pole pairs, considering the slotting effect for effect ( S Q 60= ). The slight deviation between the FEM and the analytical solution comes from the fact that the slot width and the slot opening width are not equal in the regarded FEM model (compare Fig. 5). Simultaneously, the Carter factor in Eq. (34) is derived analytically assuming straight slots [6]. Another reason for deviations between analytical and FEM results is the (very high but still) finite value of the relative permeability of iron in the FEM model, whereas ideal iron is assumed in the analytical model. As the field line length in iron parts gets shorter with higher number of pole pairs, the deviation between numerical and analytical solution decreases with increasing number of pole pairs, please compare to Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002624_jae-150151-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002624_jae-150151-Figure3-1.png",
        "caption": "Fig. 3. Physical model of the generator after the subdivision.",
        "texts": [
            " On the other hand, while the generator operates at asymmetry, the negative component appears in the stator magnetic field and the rotor magnetic field. Under the circumstance of asymmetry, there is second harmonic electromagnetic torque ripple, and it has the direct proportion with the length of the resultant positive current vector I1+, the length of the resultant negative current vector I1\u2212 and the equivalent excitation inductance Lsr of the generator. The physical model of the generator has been set up according to the actual size of the generator, moreover, control the generator to run at the off-grid condition. Figure 3 is physical model of the generator after the subdivision. Employ the transient magnetic field solver to build the finite element model. Supposing that 1) Outwards the generator stator outer circle \u0393, as shown in Fig. 3, there is no leakage flux. 2) On the condition of the rated load, the generator runs at unsaturated state. 3) The current density J , as well as the magnetic vector potential A, is the function of two-dimension coordinate (x, y). 4) The displacement current is too small when comparing with the conduction current, therefore, the displacement current can be ignored. On the condition of the no-load, to improve the velocity of the simulation, the armature stator current is compelled to be zero, additionally, the rotor windings are imposed the excitation current, then the armature voltage is within the scope of the specified permissible value"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003273_vppc.2016.7791798-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003273_vppc.2016.7791798-Figure3-1.png",
        "caption": "Fig. 3. Schematic Diagram of Vehicle Planar Dynamic Model",
        "texts": [
            ", \ud835\udc48) is the state-space representation of vehicle model, \ud835\udc47\ud835\udc51 is sample time and \ud835\udc58 is discrete time index. In this paper, the main focus about single vehicle dynamic process are on the lateral stability, vertical, roll, and pitch dynamics are therefore neglected. Meanwhile, the road is hypothesized flat, there is no slip in the longitudinal directions of tire, and the weight of vehicle is equally distributed on each wheel. The notations and coordinate system of a vehicle dynamic model are shown in Fig. 3. In Fig. 3, the coordinates of the oriented center of wheels are denoted as (\ud835\udc3f,\ud835\udc4a ), (\ud835\udc63\ud835\udc65, \ud835\udc63\ud835\udc66) are the longitudinal and lateral directions in the vehicle frame, (\ud835\udc4b,\ud835\udc4c ) are the global coordinates in the inertial frame, \ud835\udc36\ud835\udc40 is the vehicle barycenter, \ud835\udc59\ud835\udc61 is the wheel tread, \ud835\udc39\ud835\udc65\ud835\udc56 and \ud835\udc39\ud835\udc66\ud835\udc56 are the longitudinal resultant force and lateral force of tire, \ud835\udc56 = 1, 2, 3, 4, meaning the left front, left rear, right rear, and right front wheels, \ud835\udc40 and \ud835\udc3c\ud835\udc67 are, respectively, the vehicle mass and the moment of inertia, \ud835\udc59\ud835\udc53 and \ud835\udc59\ud835\udc5f are the front and rear \ud835\udc36\ud835\udc5c\ud835\udc3a distances, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001745_omae2015-41955-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001745_omae2015-41955-Figure11-1.png",
        "caption": "Figure 11. BOND GRAPH TORSIONAL MODEL OF A SHAFT SEGMENT.",
        "texts": [
            " Effort sources Se: Va, Se: Vb, and Se: Vc having sinusoidal voltages with equal amplitude but 0, -2\u03c0/3, 2\u03c0/3 phase angles, respectively have been used to excite the system. A lumped segment approach is used to model the torsional dynamics of the shafts. In the lumped segment approach, the system is divided into a number of elements, interconnected with springs [18, 21]. This model is a more cumbersome bond graph representation, and the accuracy of the model depends on the number of elements considered; however, analytic mode shapes and natural frequencies need not be determined. A total of 4 segments is used in the dynamic model for each shaft. Fig. 11 depicts the torsional dynamic sub-model for a single shaft segment. The pump impeller is modeled as a single lumped inertial body with a fluid viscous damping torque and a load torque. The damping torque and the load torque for the impellers are estimated as R\u03c9 and RL\u03c92, respectively. R and RL are viscous damping factor and impeller load factor, respectively. The bond graph dynamic model of an impeller is shown in Fig. 12. Figs. 13- 15 present the no load responses from the bond graph model of a 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001508_celc.201300256-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001508_celc.201300256-Figure1-1.png",
        "caption": "Figure 1. Diffusion domain approach employed for the simulation of the arrays of spherical nanoparticles.",
        "texts": [
            " First, the modified electrode is modelled as an array of spherical nanoparticles of the same size that are evenly distributed on the surface.[11] The conductive substrate is assumed to be large enough such that the contribution of particles situated at the edge is negligible. As a result, the problem is reduced to N identical independent problems (with N being the number of particles of the array) associated with each nanoparticle and its own diffusion space, which is approximated as cylindrical (see Figure 1). Therefore, the system can be represented in a two-dimensional cylindrical polar coordinate system [Eq. (8)]: @ci @t \u00bc Di @2ci @r2 \u00fe 1 r @2ci @z2 \u00fe @2ci @z2 i A; B\u00f0 \u00de \u00f08\u00de Regarding the boundary value problem, the initial and bulk conditions in the z direction are given by Equation (9): t \u00bc 0 t > 0; z !1 ) cA \u00bc cbulk O2 ; cB \u00bc 0mM \u00f09\u00de At the nanoparticle surface, the conditions related to the heterogeneous electrochemical and chemical processes apply [Eq. (10)]: 2014 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim ChemElectroChem 0000, 00, 1 \u2013 9 &2& These are not the final page numbers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure5-1.png",
        "caption": "Figure 5. Walking support robot.",
        "texts": [
            " (viii) The detection torque level is adjustable. (ix) The safety device consists of only passive components without actuators, controllers, or batteries. By (vii) and (ix), we can expect that the torque-based safety device prevents the robot from providing unexpected large forces to humans if the unexpected high velocity does not occur in each shaft under breaking down of the controller (Fig. 4). Furthermore, by (viii), we can adjust the detection torque level according to the requirement of each patient\u2019s gait exercise. Fig.5 shows the walking support robot with velocity and torque-based mechanical safety devices. The walking support robot has two drive units, two casters, and a force sensor. The force sensor is installed between the armrest and the body of the robot. Fig. 6 shows the drive unit. Each drive unit has a motor with an encoder and the motor torque is transmitted to Wheel via Coupling, Shaft A, Torque-based Safety Device, Shaft B, Gear 1-A, Gear 1-B, Shaft C, Gear 1-C, Gear 1-D, and Shaft D. The robot can move by controlling the two motors on the basis of the force sensor signals and the encoder signals"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.18-1.png",
        "caption": "Fig. 4.18 Translational joint connecting bodies i and j",
        "texts": [
            "60), yielding \u20acxi \u00f0nPi sin/i \u00fe gPi cos/i\u00de\u20ac/i \u00f0nPi cos/i gPi sin/i\u00de _/2 i \u20acxj \u00fe\u00f0nPj sin/j \u00fe gPj cos/j\u00de\u20ac/j \u00fe\u00f0nPj cos/j gPj sin/j\u00de _/2 j \u00bc 0 \u00f04:62\u00de \u20acyi \u00fe\u00f0nPi cos/i gPi sin/i\u00de\u20ac/i \u00f0nPi sin/i \u00fe gPi cos/i\u00de _/2 i \u20acyj \u00f0nPj cos/j gPj sin/j\u00de\u20ac/j \u00fe\u00f0nPj sin/j \u00fe gPj cos/j\u00de _/2 j \u00bc 0 \u00f04:63\u00de or in a compact form written as U\u00f0r;2\u00de q \u20acxi \u20acyi \u20ac/i \u20acxj \u20acyj \u20ac/j 8>>>>>< >>>>>>: 9>>>>>= >>>>>>; \u00bc c\u00f0r;2\u00de \u00f04:64\u00de where the vector \u03b3(r,2) is given by Eq. (4.58). The kinematic constraints and corresponding Jacobian, for velocity and acceleration equations can be obtained similarly for other types of kinematic joints. In what follows, the main kinematic variables associated with the translational joints are briefly presented. The interested reader in the detailed formulation is referred to the works by Nikravesh (1988) and Haug (1989). Figure 4.18 shows two bodies i and j connected by a translational joint, in which the slider and guide can translate with respect to each other parallel to the line of translation. Hence, no relative rotation is allowed between the bodies, and a relative translation motion in the direction perpendicular to the line of translation is not allowed either. Thus, a translational joint reduces the number of degrees-of-freedom of the system by two, which implies the need for two independent algebraic equations to represent it. A constraint equation for eliminating the relative rotation between the two bodies i and j is written as /i /j \u00f0/0 i /0 j \u00de \u00bc 0 \u00f04:65\u00de where /0 i and / 0 j are the initial rotational angles for each body. In order to eliminate the relative translation motion between the two bodies in a direction perpendicular to the line of translation, the two vectors si and d shown in Fig. 4.18 must remain parallel during the motion of the system. These vectors are defined by locating three points on the line of translation, two points on body i and one point on body j. This condition is imposed by forcing the vector product of these two vectors to remain null all the time. A simpler and alternative method consists of defining another vector ni fixed to body i and perpendicular to the line of translation having the same magnitude as si. Then, it is only required that vector d remains perpendicular to vector ni, that is nTi d \u00bc 0 \u00f04:66\u00de Therefore, Eqs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000876_ipec.2014.6869589-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000876_ipec.2014.6869589-Figure4-1.png",
        "caption": "Fig. 4. Rotor of the experimental PMSM.",
        "texts": [
            " One pole is composed of four PMs, and the thickness of four PMs which compose one pole is reduced by 10 %, 20 % and 30 % in order to imitate the demagnetization or imperfect magnetization. Fig. 3 shows the demagnetization in axial direction, we call this situation axial demagnetization hereafter. Axial length of two of four PMs is reduced by 20 %, 40 % and 60 %. Therefore, the PM volume of one pole for axial demagnetization is same as that for radial demagnetization. Nonmagnetic materials are inserted into the reduction region in order to remove eccentricity. Fig. 4 shows a photograph of the rotor of the experimental PMSM. When performing experiments with several motors, slight difference can affect the motor performance especially if the amount of demagnetization is little. In order to avoid this problem, we insert the PMs to the same rotor. Fig. 5 shows a block diagram for the PMSM controlled by V/f constant strategy. The s-function makes the reference of three stator voltages proportional to the inverter frequency f This block diagram was made on DS 11 02 DSP board, and a block DS 11 02 PWM sends a duty cycle to three-phase inverter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001690_detc2015-46173-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001690_detc2015-46173-Figure5-1.png",
        "caption": "Figure 5. Physics-based tire model using multi-layered laminated shell element",
        "texts": [
            " The number of layers, cord angles of layers, material properties are provided in each section to create the tire model data as shown :\u00a0ANCF laminated\u00a0shell :\u00a0ANSYS laminated\u00a0solid\u00a0shell :\u00a0Analytical\u00a0solution Fiber\u00a0angle\u00a0(deg) Tw is ti n g\u00a0 an gl e\u00a0 (d eg ) Figure 3. Twisting of two-layer laminate subjected to uniaxial tensile load Figure 4. Tire model creation procedure 1.\u00a0Cut\u00a0section 2.\u00a0Geometry\u00a0data\u00a0points 3.\u00a0Spline\u00a0curve 4.\u00a0Element\u00a0discretization 5.\u00a03D\u00a0tire\u00a0geometry\u00a0 6.\u00a0Tire\u00a0model\u00a0data Tire\u00a0model\u00a0 data 5 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/86618/ on 02/27/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use E in Fig. 5. The tread section consists of a carcass ply, two steel belts, a belt cover, and tread blocks. The carcass ply and steel belt are modeled as an orthotropic material with polyester and steel cords embedded in rubber, respectively. A rubber layer is considered between the upper and lower steel belts and between the carcass ply and the lower steel belt. The sidewall section is modeled by two carcass plies and a rubber that lies in between. The bead section is modeled by two carcass plies, a steel belt, and a rubber as shown in Fig. 5. Having determined the cross-section property, the three-dimensional tire geometry is generated by rotating the tire section model and the nodal coordinates are created as summarized in Fig. 4. The tire air pressure is 220 kPa that is considered by the normal distributed load applied to the inner surface of the tire. The penalty approach is used for modeling the normal contact force at each node in contact. The load-deflection curve is important for characterizing the fundamental structural properties of tires"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001462_12.2040058-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001462_12.2040058-Figure6-1.png",
        "caption": "Figure 6. Simulation of optical system.",
        "texts": [
            " Rays parallel to the axis in object space are conjugate to rays parallel to the axis in image space. The arrangement of the components of the optical zoom system is a combination of two convergent lenses and two divergent lenses, with two fixed lenses and two movable. Once a primary design was obtained, the zoom equation, which establishes the movement equation of each element to obtain the required range of magnification, was deduced. Figure 5 shows the simulation of real adaptive optical system with a perfect collimated beam, while Figure 6 shows the simulation of complete optical system with a plate beam splitter and ideal collimator and focusing lens. Proc. of SPIE Vol. 8970 89700Q-5 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 05/14/2015 Terms of Use: http://spiedl.org/terms The study of aberrations by simulation of optical system on ZEMAX (Figure 7) gave an aberration compensated result. This result is the sum of all the aberrations of each of the elements of the optical system, being the spherical aberration the predominant one"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003102_fpmc2016-1702-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003102_fpmc2016-1702-Figure4-1.png",
        "caption": "Fig. 4 Experimental device for pressure generation experiment",
        "texts": [
            " 3 Gas\u2013liquid phase change system Table 1 Characteristics of working fluid Working fluid (Chemical formula) Boiling point (1 atm)[\u2103] Heat of vaporization [kJ/kg] Coefficient of thermal expansion [\u2103-1] Fluorocarbon (C5F11NO) 50 104.65 0.00154 Water (H2O) 100 2257 0.00021 2 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90210/ on 03/06/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use To observe the generation of pressure by the GLPC, a transparent container was used to conduct a pressure generation experiment. The experimental device for the pressure generation experiment is shown in Fig. 4. In the experiment, a fixed container with a volume of 30 cm3 (cylinder wall material: acrylic, height: 42.4 mm, inner diameter: 30 mm, area: 3996.1 mm2; top and bottom surface material: aluminum, total area: 1413.7 mm2) was filled with the aforementioned fluorocarbon. The working fluid was heated and boiled by powering the constantan heater (resistance: 3.7 \u2126) at the bottom of device. In addition, a pressure sensor (SMC PSE510R06) was installed at the top of the device. By supplying a power of 331 W (voltage: 35 V, current: 9"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure2.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure2.2-1.png",
        "caption": "Fig. 2.2 (a) Configuration of mechanical tensile test. (b) Stress and strain diagram for mild steel",
        "texts": [
            " In fact, the \u201ctrue stress\u201d is considered the instantaneous direct force divided by the instantaneous crosssectional area, and the \u201ctrue strain\u201d is calculated by the instantaneous body length at any moment during elongation and the rate of increase in gauge length, i.e., \u03b5 \u00bc Z L Lo dL=L \u00bc ln L=Lo\u00f0 \u00de: \u00f02:3\u00de tested by mechanical test experiments performed on specimens with defined dimensions. The experiments are conducted in laboratories equipped with testing machines capable of loading tension or compression (Fig. 2.2a). The main mechanism for testing is pulling the sample using the machine and analyzing the change of the sample by a data-processing system. The common relationship being tested is the change in sample length under a known external force. From the results of tensile testing, a curve of direct stress against direct strain is obtained, such as the sample plot for many metals as shown in (Fig. 3.1b). In the tensile test, for example, by continuously pulling until the sample breaks, a complete tensile profile about the sample is obtained. From the origin to point A in the stress-strain diagram (Fig. 2.2b), the stress is proportional to the strain. The higher the stress, the higher the strain. The segment in this region is a straight line whose slope is called the modulus of elasticity. This line is no longer linear when the stress exceeds point A, and therefore, this point is defined as an elastic limit which is the upper limit to the linear line. This sweeping idealization and generalization applicable to all materials is known as Hooke\u2019s law. 2.2 Solid Properties 35 Hooke\u2019s law is the principle that elastic materials obey when they are released from stress and relax back into their original shapes",
            " This stiffness equation can apply to both compression and extension cases, but the deformation (x) is positive at extension and negative at compression (this property is the same with force). Looking back to stress-strain perspective, the size can be eliminated by using stress and strain instead of force and deformation. Recall that in Eqs. 3.1 and 3.2, we obtain a relation k \u00bc F/x \u00bc \u03c3A/(\u03b5L), and hence \u03c3 \u00bc FL Ax \u03b5 \u00bc E\u03b5 \u00f02:4\u00de where E is called the modulus of elasticity (or elastic modulus). The slope of the straight line in the elastic region of the stress-strain curve (Fig. 2.2b) is E. As the strain is dimensionless, as mentioned before, the unit of E is the same as the unit of stress. A slight increase in stress above the elastic limit will result in breakdown of the material and cause it to deform permanently; this process is called yielding. The stress at yielding is called the yield stress. Once this yield point is reached, it delimits the elasticity from the plastic region. Beyond this point, plastic deformation begins, and the material in this state becomes perfectly plastic in which the specimen will elongate (strain) without any increase in load",
            "3a (point B), it is located at the line offset where an arbitrary amount of 0.2% of strain is drawn parallel to the straight-line portion of the stress-strain diagram. Furthermore, the area enclosed by the loop in Fig. 2.3b corresponds to dissipated energy released through heat. Perfect elasticity of materials (between points B and C) occurs when there is no dissipation of any energy for deformations under a monotonic or cyclic loading. 36 2 Basic Material Properties As the specimen is subjected to an increasing load beyond the point C in Fig. 2.2b, the curve rises continuously before becoming flat when reaching the ultimate stress, which is the maximum force that the specimen can endure before causing it to break or carry less load. The rising of the curve between points C and D is called strain hardening, which occurs after the yielding ends when any further load is applied to the specimen. In strain hardening, the material crystalline structure changes as many crystalline dislocations propagate and interact with each other, resulting in the increased resistance of the material to further deformation",
            " The ultimate stress or strength is the maximum stress level reached in the mechanical test (point D in the stress-strain curve). Beyond the strain at the ultimate stress, a particular cross-section of the specimen will have a rapid reduction in its area as shown in Fig. 2.4. As a result, a \u201cneck\u201d along the specimen tends to form; and such necking behavior can be observed for some metals. As the cross-sectional area of the neck gradually decreases, the smaller the area is, the less the load that this portion of specimen can carry. Therefore, the stress-strain curve goes downward (between points D and E in Fig. 2.2b) in the necking region, and, eventually, the specimen reaches its maximum possible strain and breaks at the fracture strain (point E). Considering that the instantaneous load in tension is given by F \u00bc \u03c3A for a material in the strain hardening region during a tensile test. The criterion for the instability in tensile test (necking) can be formulated as the condition that F is just beyond the maximum (i.e., \u03c3 or \u03c3 is just beyond the strength), meaning that dF/ d\u03b5 0, where \u03b5 is the flow strain. Near but slightly before reaching the maximum load, the uniform deformation conditions can be assumed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002751_b978-0-12-404616-0.00004-9-Figure4.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002751_b978-0-12-404616-0.00004-9-Figure4.10-1.png",
        "caption": "FIGURE 4.10",
        "texts": [
            " There are several other advanced types of shock absorber based on electro-rheological fluids or magneto-rheological fluids which use, respectively, an electrical field or a magnetic field to change the viscosity of the hydraulic fluid. It is likely that one or other of these will be more widely adopted in the years ahead so as to give the smoothest possible ride. Modern heavy trucks generally have air suspensions with hydraulic dampers fitted for safe handling and an even ride. The wheel is attached to the body of the vehicle by means of metal struts. Two configurations are widely adopted \u2013 namely, the double-wishbone strut and the MacPherson strut; these are shown schematically in Figure 4.10(a) and Figure 4.10(b), respectively. Each version has its advantages and limitations. Double-wishbone struts were once common on cars, but are now mostly used on trucks and larger vehicles. The system has two sets of transverse links that connect the top and bottom extremities of individual wheel carriers to the structure. The suspension spring is usually a steel coil clamped between the body structure of the car at one end and one of the wishbone members at the other. In addition, there will be a telescopic hydraulic damper which is very resistant to large, abrupt movements but quite compliant to small, slow ones"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003362_cgncc.2016.7828902-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003362_cgncc.2016.7828902-Figure2-1.png",
        "caption": "Figure 2. Echelon formation configuration diagram.",
        "texts": [],
        "surrounding_texts": [
            "Basic task of jamming is to hinder normal work of radar, and to weaken its detection performance so that it would fail to detect effectively. Through the design of spatial location, we can optimize the jamming effect. Steps of jamming effectiveness modeling proposed are as follows: First, get radar interference signal power at the receiving end in the implementation of jamming according to the signal detection theory. Second, get the suppression sector via the boundary of effective jamming area\u2019s curve equation. Third, merge attack range evaluation, distance loss assessment and interference range into a single jamming effectiveness evaluation model of cooperative formation: B D SC B D S\u03c9 \u03c9 \u03c9= \u00d7 + \u00d7 + \u00d7 . (2) B is attack range evaluation. D is distance loss assessment. S is interference distance range. B\u03c9 , D\u03c9 , S\u03c9 are weight coefficients, these three parameters depend on the degree that decision maker emphasize each effectiveness."
        ]
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.14-1.png",
        "caption": "FIGURE 6.14",
        "texts": [
            " At a given roll angle for the lumped mass model the displacement and hence the force in the spring will be too large when compared with the corresponding situation in the linkage model. For the swing arm model the instant centre about which the suspension pivots is often on the other side of the vehicle. In this case the displacement in the spring is approximately the same as at the wheel and a similar problem occurs as with the lumped mass model. For all three simplified models this problem can be overcome as shown in Figure 6.14 by using an \u2018equivalent\u2019 spring that acts at the wheel centre. As an approximation, ignoring exact suspension geometry, the expression (Eqn 6.2) can be used to represent the stiffness, kw, of the equivalent spring at the wheel kw = Fw/\u03b4w = (Ls/Lw) Fs / (Lw/Ls) \u03b4s = (Ls/Lw)2 ks \u00f06:2\u00de The presence of a square function in the ratio can be considered a combination of both the extra mechanical advantage in moving the definition of spring stiffness to the wheel centre and the extra spring deflection at the wheel centre"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.4-1.png",
        "caption": "Fig. 3.4 Completely closed kinematic chain (kinematic modeling)",
        "texts": [
            " A system with kinematic loops forms a partially closed kinematic chain, when \u2022 single partial systems form open chains or \u2022 multiple closed partial systems are connected to each other in an open chain (Fig. 3.3). A chain can be considered completely closed when \u2022 each body is a part of a multibody loop and \u2022 each loop has at least one body that is connected to another loop. A mechanism, by definition, must be a partially or completely closed kinematic chain (for more details see Sect. 3.4.1) (Fig. 3.4). Kinematic chains can be grouped into three distinct motion categories: Planar kinematic chains In a planar kinematic chain all body points move inside or parallel to a reference motion plane (Fig. 3.5). Because of this, the motion of each body in the system has one rotational and two translational motion components. Relative motions between joints must be either translational displacements, that are parallel to the reference motion plane, or rotations, that are normal to the motion plane. Spherical kinematic chains In a spherical kinematic chain, all body points move on concentric spherical surfaces around a fixed point O in the center (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002624_jae-150151-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002624_jae-150151-Figure2-1.png",
        "caption": "Fig. 2. Projections of the stator and rotor positive and negative current vector on d-axis and q-axis.",
        "texts": [
            " Here, the stator positive and negative current vector component on d-axis are i1sd+, i1sd\u2212 respectively, the stator positive and negative current vector component on q-axis are i1sq+, i1sq\u2212 respectively, the rotor positive and negative current vector component on d-axis are i1rd+, i1rd\u2212 respectively, and the rotor positive and negative current vector component on q-axis are i1rq+, i1rq\u2212 respectively. The projections of the stator and rotor positive and negative current vector on d-axis and q-axis are shown in Fig. 2. Their relationship is written by the Eq. (9).\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 i1s+ = i1sd+ + i1sq+ i1s\u2212 = i1sd\u2212 + i1sq\u2212 i1r+ = i1rd+ + i1rq+ i1r\u2212 = i1rd\u2212 + i1rq\u2212 (9) The period of the vector motion in Fig. 2 is T1. Then, the theorem of the vector operation harnessed, the average electromagnetic torque in one period can be given by Te = \u222b T1 0 pLsr( i1r+ + i1r\u2212)\u00d7 ( i1s+ + i1s\u2212)dt T1 = pLsr \u222b T1 0 ( i1r+ \u00d7 i1s+ + i1r+ \u00d7 i1s\u2212 + i1r\u2212 \u00d7 i1s+ + i1r\u2212 \u00d7 i1s\u2212)dt T1 (10) It is considered that i1r+ i1s+ i1r\u2212 i1s\u2212, so i1r\u2212 i1s\u2212 can be neglect. Bring the Eq. (9) to (10). It is written by pLsr \u222b T1 0 ( i1r+ \u00d7 i1s+ + i1r+ \u00d7 i1s\u2212 + i1r\u2212 \u00d7 i1s+ + i1r\u2212 \u00d7 i1s\u2212)dt T1 = pLsr [ I1r+I1s+ sin \u03b8r + \u222b T1 0 ( i1r+ \u00d7 i1s\u2212 + i1r\u2212 \u00d7 i1s+)dt T1 ] (11) = pLsr [ I1r+I1sq+ + \u222b T1 0 ( i1rd+ \u00d7 i1sq\u2212 + i1rq+ \u00d7 i1sd\u2212 + i1rd\u2212 \u00d7 i1sq+ + i1rq\u2212 \u00d7 i1sd+)dt T1 ] The positive components on d-axis and q-axis spin with d-axis and q-axis at the synchronous speed, however, the negative components on d-axis and q-axis spins off d-axis and q-axis at the twice synchronous speed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002263_meacs.2014.6986869-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002263_meacs.2014.6986869-Figure4-1.png",
        "caption": "Fig. 4. Stand for registration of vibrosignals",
        "texts": [
            " This database contains records of vibration signals of rolling bearings, containing defects (on outside and inside track, and on the rolling body (Fig. 3), and without them. The sizes of defects are 0.007, 0.014, 0.021, and 0.028 inches in diameter. The defects of the outer track are stationary, therefore, the position of the defect relatively the load zone of the bearing influences the vibration signal, generated by the bearing. To investigate the influence of this effect, the defects were applied on the outer track in three positions. The scheme of the stand to record vibration signals is shown in Fig. 4. The test bearing is fixed on the motor shaft. Accelerometers are used to register the vibration signal, which are placed on the engine case and at the place of load connection (DE) , next to the fan (FE) and on the supporting device (BA) . Vibrosignals' recording is performed with a sampling rate of 12 kHz and 48 kHz. The results of testing of the eXlstmg classification methods of the rolling bearing status are shown in Table 1. Table 2 shows the results of the research of the proposed method according to vibration signals which were recorded in different parts of the stand"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001968_s40435-016-0225-2-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001968_s40435-016-0225-2-Figure2-1.png",
        "caption": "Fig. 2 Exerting the controlling torques on biped robot",
        "texts": [
            " (2) and (3), the transition equations that relate the state variables before and after the collision are obtained as follows: \u23a7 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03b8 \u03b8\u0307 \u03d5 \u03d5\u0307 \u03b1 \u03b1\u0307 \u23ab \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ad + = A (\u03b8, \u03b1) \u23a7 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03b8 \u03b8\u0307 \u03d5 \u03d5\u0307 \u03b1 \u03b1\u0307 \u23ab \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ad \u2212 , (4) where the matrix A is defined as below: A= \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 \u22121 0 0 0 0 0 0 \u03bc (cos(2\u03b8)\u2212cos (2\u03b1))+2 cos (2\u03b8) 2\u03b2 sin2 (2\u03b8)+2\u03bc sin2 (\u03b8\u2212\u03b1)+2 0 0 0 0 2 0 0 0 0 0 0 (1 \u2212 cos(2\u03b8)) \u03bc (cos (2\u03b8)\u2212cos (2\u03b1))+2 cos (2\u03b8) 2\u03b2 sin2 (2\u03b8)+2\u03bc sin2 (\u03b8\u2212\u03b1)+2 0 0 0 0 0 0 0 0 1 0 0 \u2212 cos (\u03b8\u2212\u03b1) \u03ba \u03bc (cos (2\u03b8)\u2212cos (2\u03b1))+2 cos (2\u03b8) 2\u03b2 sin2 (2\u03b8)+2\u03bc sin2 (\u03b8\u2212\u03b1)+2 + cos (\u03b8+\u03b1) \u03ba 0 0 0 1 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 (5) In this section, first a biped robot model without the upper body, the simplest model of biped robot proposed by Garcia et al. [6], is used as shown in Fig. 2. Also in this model, the transient motion is damped and a steady cycle is reached before the collision.Therefore, the system loses someamount of its potential energy in every step. This decrease in potential energy causes the kinetic energy of system to increase in the middle of a gait. In a stable gait, this increase is equal to the amount of energy that is lost by collision of the leg to the ground. If at the end of each step, the reference line is set on the collision point, the sum of mechanical energy of the system seems to be constant; regardless of the robot is descending down the ramp",
            " This issue beside the fact that two stable passive limit cycles cannot be existed simultaneously on one ramp indicates that there are very strong inherent dynamical properties in the robot that tend it to choose a specific passive gait cycle. Onemethod of exerting energy to the system in order to reach the preferred normalized energy level is to apply torques on the robot joints. These torques can be exerted on the contact point of the stance leg with the ground or the swing leg in the case of simplest model. The methods of exerting torques are shown in Fig. 2. In this figure, TH is the applied torque on swing leg and TA is the applied torque on the contact point of the stance leg with the ground. In order to consider these torques in the dynamical equations of the system, it is enough to add the generalized forces at the left side of Lagrange equations [36]. By scaling the time, i.e. multiplying both sides of equation by l g , and then subtracting the second equation from the first one, one has: [ \u03b2 (1 \u2212 cos\u03c6) + 1 \u2212\u03b2 cos\u03c6 (1 \u2212 cos\u03c6) 1 ] [ \u03b8\u0308 \u03c6\u0308 ] + sin \u03c6 [ \u03b2 (\u03b8\u0307 + \u03c6\u0307)2 \u2212\u03b8\u03072 ] + [\u2212(1 + \u03b2) sin (\u03b8 + \u03b3 ) sin (\u03c6 + \u03b8 + \u03b3 ) ] = [ uA \u2212 uH 1 \u03b2 uH ] = [ 1 \u22121 0 1 \u03b2 ] [ uA uH ] (8) where the non-dimensional torques uH and uA are defined as below: { uH = 1 Mlg TH uA = 1 Mlg TA (9) Since TH and TA are general forces in the general coordinates \u03d5 and \u03b8 , the amount of changes in the energy of system in every moment can be found as belowwith regard to Fig. 2: By normalizing t in Eq. (10), i.e. converting t into \u221a l g t , and dividing the equation by Ml2, one has: E\u0307N = uA\u03b8\u0307 + uH \u03d5\u0307 = \u0307TU, (11) where vectors U and are defined as: U = { uH uA } , = { \u03d5 \u03b8 } . (12) Therefore, in order to converge the normalized energy of robot EN to the normalized target energy Etar N exponentially, it is enough to satisfying the following equation: E\u0307N = \u2212\u03bb(EN \u2212 Etar N ). (13) Comparing Eqs. (11) and (13), the controlling inputU for converging EN to Etar N exponentially is obtained as below: \u0307TU = \u2212\u03bb (EN \u2212 Etar N )",
            " Therefore, it can be said that in order to create a proper energy pass tracking for the system with upper body movements, it is enough that the value of\u2212\u03b7(cos (\u03b8 \u2212\u03b1) \u03b1\u0308+ sin (\u03b8 \u2212\u03b1) \u03b1\u03072) be exactly equal to the proper uA value which produces the same energy pass tracking in the system. The proper uA for tracking a special energy pass is defined in Eq. (15). Thus the proper upper bodymovements can be obtained by solving the following equation: \u2212\u03b7 ( cos (\u03b8\u2212\u03b1) \u03b1\u0308+sin (\u03b8\u2212\u03b1) \u03b1\u03072 ) = \u2212\u03bb E \u2212 Etar \u03b8\u0307 , (17) where E and Etar are the normalized energy of the simplest biped robot model shown in Fig. 2, i.e. the model without upper body, and the normalized target energy, respectively. There exist two singular points in Eq. (17). Similar to Sect. 3.1.1, the first point is \u03b8\u0307 = 0 where the control moment converges to infinity and the appliedmoment should bemodified by a saturation level. Second point is where cos (\u03b8 \u2212 \u03b1) is equal to zero and therefore a normal singularity occurs. In this situation, Eq. (17) will be turned into a first order equation from a second order one. Therefore, the equation of the proper upper body movements for desired energy tracking will be modified as below: \u23a7 \u23aa\u23a8 \u23aa\u23a9 cos (\u03b8\u2212\u03b1) \u03b1\u0308+sin (\u03b8\u2212\u03b1) \u03b1\u03072= \u03bb \u03b7 E\u2212Etar \u03b8\u0307 \u2223\u2223\u03b8\u0307 \u2223\u2223>\u03b5, \u03b8\u2212\u03b1 = \u03c0 2 cos (\u03b8 \u2212 \u03b1) \u03b1\u0308 + sin (\u03b8 \u2212 \u03b1) \u03b1\u03072 = \u03bb \u03b7 E\u2212Etar \u03b5 \u2223\u2223\u03b8\u0307 \u2223\u2223 < \u03b5 \u03b1\u03072 = \u03bb \u03b7 E\u2212Etar \u03b8\u0307 \u03b8 \u2212 \u03b1 = \u03c0 2 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000434_j.proeng.2015.12.473-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000434_j.proeng.2015.12.473-Figure3-1.png",
        "caption": "Fig. 3. Integration of the project with other disciplines. Finite elements analysis of the static problem [10]",
        "texts": [],
        "surrounding_texts": [
            "Work between different areas of expertise on the considered problem can improve the communication of transversal knowledge in engineering problems, especially in manufacturing technologies. This experience has allowed the authors to consolidate a workgroup and to start the creations of what is called \u201clearning objects\u201d as an instrument for teachers in their respective subjects. The interrelations of solutions are enhanced even with other areas of expertise different from the workgroup [10]. The material is mainly aimed to seminars and practice where the students can promote the discussion of different solutions."
        ]
    },
    {
        "image_filename": "designv11_64_0003051_0954407016629517-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003051_0954407016629517-Figure2-1.png",
        "caption": "Figure 2. Dependences between the steering angles and the slip angles of (a) the front axle and (b) the rear axle.",
        "texts": [
            " When this simplification is taken into account, the motion of the vehicle can be described by the system of equations m _x _c + \u20acy + Y1 + Y2 +Fy = 0 \u00f03\u00de Jz\u20acc + Y1l1 Y2l2 + Mz = 0 \u00f04\u00de In the case of steady-state cornering, it can be assumed that the lateral forces Y1 and Y2 are functions of the axle slip angles a1 and a2. The system of equations presented above can be converted to the form \u20acy = Y1 a1\u00f0 \u00de+Y2 a2\u00f0 \u00de m _x _c+Fy m \u00f05\u00de \u20acc = Y1 a1\u00f0 \u00de l1 Y2 a2\u00f0 \u00del2 +Mz Jz \u00f06\u00de The dependences between the steering angles and the slip angles of the rear axle and the front axle are shown in Figure 2. They can be described by a1 =arctan v1y vx d1 \u00f07\u00de at University College London on June 5, 2016pid.sagepub.comDownloaded from a2 =arctan v2y vx + d2 \u00f08\u00de where v1y is the lateral velocity of the front axle and v2y is the lateral velocity of the rear axle. The longitudinal velocity vx (= _x), the steering-wheel d1 of the front angle and optionally the steering-wheel angle d2 of the rear axle (usually d2 = 0) of the vehicle, together with the disturbing forces denoted by Fy and Mz (e.g. the side wind), are the input signals of the vehicle model described by the system of equations (5) and (6)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003069_cp.2014.1237-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003069_cp.2014.1237-Figure1-1.png",
        "caption": "Figure 1. Circular whirling orbit and coordinate systems.",
        "texts": [
            " All the CFD simulations here are aimed to validate reliability of the CFO method and study its effect factors to make it have a better approaching towards the experimental results. The numerical simulations are performed using the same annular seal measured by Kanemori [\\8] in 1992 in order to make the comparisons. The seal is a liquid annular seal with large aspect ratio. Its main parameters are listed in Table 1. The fluid forces induced by the whirling motion of rotor around the centre of stator were measured by Ka nemori under different rotating speeds and different pres sure differences. Figure 1 shows the whirling orbit, coor dinate systems and description styles of fluid forces adopted by Kanemori [ 18] and they are still employed in present paper. In Figure 1, e represents the whirling ra dius listed in Table 1; m is the rotating speed of rotor and o is the whirling speed around the centre of stator. The circular whirling orbit can be expressed in Equation ( 1). {X = ecos (nt) (I) y = esin (nt) Considering computational amount and computing time, only one operation condition, rotating speed 1080 rpm and pressure different 907 KPa, is numerically simulated in this paper to discuss the accuracy and its effect factors of CFO method. The temperature of water is assumed 40\"C for not being given in ref",
            " As can be seen in Figure 3 and Figure 4, the forces vary moderately around the Realizable k-8 (y+ < 3) line with the radial grid densities and wall y+ changing and terrible results don't emerge, indicating insensitivity of the enhance wall function on wall y+ and its suitability for the fluid force computation in annular seal. 3.2 Quasi-Steady Whirling Model [n present study, a displacement e (0.049 mm) along the +X axial direction is imposed on the rotor to generate the eccentric numerical model. By solving the eccentric flow field under the dynamic reference frame attached to the whirling rotor, the transient whirling problem shown in Figure 1 can be converted into a steady problem. From the steady simulation, the fluid forces in rectangular coor dinate system, Fx , Fy, can be obtained by the integra- tion of pressure on the rotor surface. According to Figure 1, the radial and tangential forces can be expressed in Equation (2). Fr = Fx, Ft = Fy (2) Then, by performing several simulations under different whirling speeds, the fluid forces under different whirling speeds can be obtained. 3.3 Transient Whirling Model To perform the full transient simulation, two obstacles need to be settled: one is moving mesh problem resulting from rotor whirling motion; the other is the imposition of rotating speed on the rotor surface as the centre of rotor is not stationary and keeps whirling around the centre of stator",
            " The imposition of rotating speed on rotor surface is achieved also by a user-defmed subrou tine based on the DEF[NE PROF[LE macro in Fluent which can assign the velocity components for each grid face on rotor surface at each time step. [n present simulation, the time of rotor whirling one de gree is chosen as the time step. And, first order implicit formulation is used for the transient term discretization. By the integration of pressure on rotor surface, the fluid forces (Fx, Fy) at each time step can be obtained. Ac cording to Figure 1, the radial and tangential components of fluid forces can be expressed in Equation (3). [Fr ] [COS(nt) sin(nt) ][FX] Ft - -sin(nt) cos(nt) Fy (3) Substituting Fx. Fy in Equation (3), the radial and tangential forces (Fr, Ft) on rotor under specific whirling motion can be obtained. Likewise, by performing several transient simulations under different whirling speeds, the reaction forces under different whirling speeds can be obtained. 3.4 Obtaining Force Rotordynamic Coefficients For the whirling motion with eccentricity 12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.48-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.48-1.png",
        "caption": "Fig. 10.48 Coupled inductor minor B\u2013H loop",
        "texts": [
            " The inductors of the SEPIC and C\u0301uk converters, as well as of multiple-output buck-derived converters and some other converters, can be coupled. The inductor current ripples can be controlled by control of the winding leakage inductances [97, 98]. Dc currents flow in each winding as illustrated in Fig. 10.47b, and the net magnetization of the core is proportional to the sum of the winding ampere-turns: Hc(t) = n1i1(t) + n2i2(t) c Rc Rc +Rg (10.105) As in the case of the single winding filter inductor, the size of the minor B\u2013H loop is proportional to the total current ripple (Fig. 10.48). Small ripple implies small core loss, as well as small proximity loss. An air gap is employed, and the maximum flux density is typically limited by saturation. As discussed in Chap. 6, the flyback transformer functions as an inductor with two windings. The primary winding is used during the transistor conduction interval, and the secondary is used during the diode conduction interval. A flyback converter is illustrated in Fig. 10.49a, with the flyback transformer modeled as a magnetizing inductance in parallel with an ideal transformer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003137_0954410016676847-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003137_0954410016676847-Figure1-1.png",
        "caption": "Figure 1. Lambert\u2019s problem with respect to the geocentric-equatorial reference frame.",
        "texts": [
            "14 This new method employs a higherorder Be\u0301zier implicit function and a more reliable and effective update scheme compared with Mortari\u2019s method, taking full advantage of the monotonicity and boundedness of the independent variable in an appropriate formulation of Lambert\u2019s problem. No initial guess is required for this new algorithm, and thus the common failure caused by the poor initial guess in most of the existing Lambert solvers is addressed. With the new formulation and numerical method, numerical testing and comparison is provided to demonstrate the robustness and the notable improvement in computational efficiency in solving Lambert\u2019s problem. The geocentric-equatorial inertial (GEI) reference frame (X, Y, Z) is used to describe Lambert\u2019s problem, as shown in Figure 1. The (X, Y)-plane lies in the Earth\u2019s equatorial plane and the X-axis points toward the vernal equinox. Denote the initial and target points by P1 and P2 and the corresponding radial positions by r1 and r2. The problem is to determine the conic orbit for an orbiting body which starts from P1 at time t1 and must arrive at P2 at the specified time t2. For a two-body orbit, there are six integration constants, referred to as the orbital elements, including the semi-major axis a, eccentricity e, inclination of the orbital plane i, right ascension longitude of the ascending node , argument of periapsis !, and true anomaly f. The angles , i and ! specify the orientation of the orbital plane with respect to the GEI frame.15 The elements a and e determine the size and shape of the elliptic orbit respectively, and the true anomaly f relates each position in the orbit to a time. If the transfer time t\u00bc t2 t1 from P1 to P2 is given, then Lambert\u2019s problem is to find the conic orbit joining P1 to P2, which is completely determined by the six orbit elements. As illustrated in Figure 1, the angles and i can be determined by the unit vector of angular momentum 1h and the unit vector of Z-axis 1Z, that is, cos\u00f0 \u00de \u00bc 1X\u00f01Z 1h\u00de \u00f01\u00de cos\u00f0i\u00de \u00bc 1h1Z \u00f02\u00de where 1X\u00bc [1, 0, 0], 1Z\u00bc [0, 0, 1] and 1h \u00bc r1 r2= k r1 r2 k. at UNIV OF WESTERN ONTARIO on November 25, 2016pig.sagepub.comDownloaded from Once and i are found, the remaining four elements can be determined within the orbital plane. For convenience of description, a two-dimensional reference frame in the orbital plane is introduced. As shown in Figure 2, the x-axis is along the node line, which is the intersection of the orbital plane and the equatorial plane, and the y-axis is perpendicular to the x-axis and inside the orbital plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001377_s00542-014-2085-z-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001377_s00542-014-2085-z-Figure10-1.png",
        "caption": "Fig. 10 Displacement simulation model",
        "texts": [
            " The relationship between the force on the TFPM and the current input into the electromagnet is linear, with a value of 37.8 mn/a, measured by Zhi et al. (2013). Using the above parameter, the experimental data in Fig. 11a was converted into a relationship between force and displacement, shown as the solid line in Fig. 11b. 1 3 The diaphragm displacement was also calculated using a large deformation analysis model by 3D caD (SolidWorks Simulation, SolidWorks corp). In the simulation model, the edge of the PDMS film is clamped and a uniform pressure is distributed in the central circular area, as shown in Fig. 10. The simulated result is compared with the experimental one in Fig. 11b, which shows there is good agreement between the simulated and experimental data, thus demonstrating that the Young\u2019s modulus of 2 MPa and Poisson\u2019s ratio of 0.45 for the PDMS used in the simulation are the same as the conventional values (Suzuki et al. 2011; lu and Zheng 2004). Thus, we can conclude that the PDMS retains its mechanical properties after XeF2 gas etching. For the magnetic diaphragm fabricated in Fig. 4b, according to the results of the evaluation, the TFPM retained its magnetic performance (recovered after re-magnetization); meanwhile, the PDMS layer retained its mechanical performance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure8.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure8.6-1.png",
        "caption": "Fig. 8.6 Working principle of laser surface cladding",
        "texts": [
            " Using this procedure, one can maintain homogeneous properties on the bulk body due to little thermal penetration, resulting in little distortion, smooth surfaces, and reduced work after processing. As other laser processes, this meltand-blow process can be automated, and hence, laser surface melting is of industrial interest for materials such as cast irons, stainless steels, and titanium. Laser cladding is a surface processing technique used for adding one material to the surface of another in a controlled manner that will produce a coating with 0.3\u20131 mm thick by fusion bonding. It generates a clad track by inputting powder particles to a molten pool made by moving laser (Fig. 8.6). The adding powder material can be the same or different from the bulk sample material. This enables the applied material to be deposited selectively wherever it is required. Nowadays, there are still demands for improvements on surface wear resistance. Conventional manufacturing processes such as press molds, stamping tools, casting forms, shafts, and other machinery elements usually come with the process of surface wear which significantly affects the lifetime of cutting tools and machines. Laser cladding is one of the most effective ways to improve surface characteristics, including corrosion resistance, wear resistance, and heat resistance",
            " Developed based on the concept of rapid prototyping, laser sintering is a method that uses laser and powder to generate a three-dimensional model which has the benefit of not requiring a supporting material. Laser sintering utilizes laser power as a heat source on the traditional powder compact sintering technology. This technique has unique advantages that are not easily achieved by conventional sintering furnaces. Due to a laser beam\u2019s concentration of heat and penetration, suitable for small areas, thin products can be sintered. It is easy to sinter powder or flake compacts different from the matrix composition. The typical process is illustrated in Fig. 8.6. Briefly, this process is based on assembling materials by heating microbeads of known materials to their melting temperature in order to have the microbeads attach together into the defined product geometry. Laser sintering can use nearly all kinds of material as a \u201cbase,\u201d no matter metal, polymer, alloy, or ceramics as products with very flexible geometry. This technology generates 3D parts by selectively fusing thermoplastic, ceramic, or metallic powders with the localized heat from an infrared laser, usually a CO2 232 8 Laser Metal Processing laser"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001174_0954405414554016-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001174_0954405414554016-Figure6-1.png",
        "caption": "Figure 6. Contact model of the rolling body of an angle-contact ball bearing: (a) dimensions of angle-contact ball bearing and (b) contact deformation of rolling body.",
        "texts": [
            " The cage is an important component of a rolling bearing to keep and warranty an adequate position of its rolling bodies, but it has little impact for contact stiffness between rolling and rings, so the cage is neglected in model for convenience and the rolling bodies\u2019 position is guaranteed by several springs which can only be compressed. While working, a rolling body produces deformations that link the inner and outer rings, which can be viewed as two springs in series. The radial components of the contact stiffness of all rolling bodies are parallel, as are the axial components. Contact stiffness of a single rolling body. Dimensions related to the angle-contact ball bearing and the deformations between a rolling body and an inner and outer ring are shown in Figure 6, and the corresponding parameters are defined in Table 3. According to the Hertz contact theory, the contact stiffnesses of the inner and outer rings are calculated from equations (1)\u2013(3) (i: inner ring and o: outer ring) (Table 4). Ri1 = Db 2 , R0i1 = fiDb Ri2 = Db 2 , R0i2 = Di +Do 2Db cos a0 4 cos a0 ai = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ri1R 0 i1 Ri1 R0i1 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Rim 3 r bi = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ri2R 0 i2 Ri2 +R0i2 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Rin 3 r Ri = ffiffiffiffiffiffiffiffiffiffiffiffiffiffi RimRin p aibi =Ridi Ki = 4 3 ER 1=2 i d 1=2 i = 16EF2R2 i 9 1=3 or Kic = 4 3 ER 1=2 i \u00f021\u00de Ro1 = Db 2 , R0o1 = foDb Ro2 = Db 2 , R0o2 = Di +Do +2Db cos a0 4 cos a0 ao = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ro1R 0 o1 R0o1 Ro1 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Rom 3 r bo = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ro2R 0 o2 R0o2 Ro2 3 s = ffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 4 F E Ron 3 r Ro = ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi RomRon p aobo =Rodo Ko = 4 3 ER1=2 o d1=2 o = 16EF2R2 o 9 1=3 or Koc = 4 3 ER1=2 o \u00f022\u00de at NANYANG TECH UNIV LIBRARY on June 5, 2016pib"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002590_s11071-016-2917-8-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002590_s11071-016-2917-8-Figure4-1.png",
        "caption": "Fig. 4 2-DOF model of the drivetrain vibration",
        "texts": [
            " With this limit, the first equation in (4) is reduced to q\u03080 = 0, which implies the algebraic relation P = Tds. The second and third equations are reduced to the following forms: { \u03b8\u03081 = \u03c11(Ttm \u2212 Tds), \u03b8\u03082 = \u03c11Tds \u2212 (\u03c11 + \u03c12) Ttm + \u03c12Teg, (5) where \u03c1i := (Ii )\u22121 is the inverse of the corresponding inertia moment. These describe vibrations of the relative angles \u03b81 and \u03b82 independently of q0. Therefore, we have obtained a 2-DOF model (5) of the torsional vibration of the drivetrain, whose mechanical description is given in Fig. 4. 3.3 Modeling of torque transfer characteristics Focusing on the torque transfer characteristics from Ttm to Tds, we develop a closed-form expression of Tds as a function of \u03b81 and \u03b8\u03071: Tds = F(\u03b81, \u03b8\u03071). (6) To specify F(\u03b81, \u03b8\u03071), we suppose that the stationary intervals in Fig. 2 are induced by a total backlash of gears and couplings between the rotor R1 and the fixed base, or by a clearance-type nonlinearity with respect to \u03b81. We also suppose that the clearance induces a torsional impact behavior of rotor R1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.18-1.png",
        "caption": "Fig. 35.18 Single row thrust ball bearing.",
        "texts": [],
        "surrounding_texts": [
            "In a taper roller bearing the line of action of the resultant load through the rollers forms an angle with the bearing axis. Taper roller bearings are therefore particularly suitable for carrying combined radial and axial loads. The bearings are of separable design, i.e. the outer ring (cup) and the inner ring with cage and roller assembly (cone) may be mounted separately. Single row taper roller bearings can carry axial loads in one direction only. A radial load imposed on the bearing gives rise to an induced axial load which must be counteracted and the bearing is therefore generally adjusted against a second bearing. Two and four row taper roller bearings are also made for applications such as rolling mills."
        ]
    },
    {
        "image_filename": "designv11_64_0000072_978-1-4614-9126-2-Figure6.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000072_978-1-4614-9126-2-Figure6.4-1.png",
        "caption": "Fig. 6.4 (a) A cylinderpiston representing a closed system (b) A bioreactor representing an open system",
        "texts": [
            " 6.2). Here, a sphere comes to rest at the bottom of a curved surface. At this point, the forces acting on it are in equilibrium. (iii) State: physical condition of a system that can be described by specifying a limited number of observable variables (Fig. 6.3). Definitions.We call a system a sector of the universe that is delimited and that you wish to study. The remaining part of the universe is the surroundings. A closed system is one that does not exchange matter with the surrounding area (Fig. 6.4a). When there is exchange of matter between a system and its surroundings, it is considered an open system (Fig. 6.4b). An isolated system is one that does not exchange matter or energy with its surroundings. A process corresponds to a sequence of transformations, quantified by changes in the properties that describe a system, that can get you from one state or initial condition to another state or end condition through one or several stages (Fig. 6.5). First, let us clarify that heat and temperature are concepts that in colloquial language are normally confused. For example, the expression \u201cIt\u2019s hot\u201d is commonly used to indicate that the temperature is high, although the word heat is used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000433_j.proeng.2015.07.170-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000433_j.proeng.2015.07.170-Figure1-1.png",
        "caption": "Figure 1: Mesh of blade \u2013 shaft interface Figure 2: Experimental setup to measure puck stiffness",
        "texts": [
            " Due to the complex shape of the taper at the hosel and curve of the blade, the geometry of the stick was modeled using Creo Parametric (PTC, Needham, MA). The solid model was then imported into LS-PrePost (Livermore Software Technology Corporation, Livermore, CA) where the shaft and blade were meshed using 8-node solid brick elements. The mesh between the wood shaft and fiberglass blade was modeled to resemble the geometry seen in the stick by morphing the shapes of the elements at the interface as seen in Figure 1. The stick had a total of 3528 elements, 2868 elements for the shaft of the stick and 660 elements for the blade. The shaft had a rectangular cross section of 20 mm by 30 mm and the blade had a rectangular cross section of 65 mm by 5.6 mm at its centre. This corresponded to a second moment of inertia of 20,936 mm4 and 1034 mm4 for the stick and blade (measured about the compliant axis), respectively. Both the blade and shaft were modelled as elastic, isotropic materials. Since the stick and blade primarily undergo simple bending during puck impact, the isotropic model used here is a good approximation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000913_j.mechmachtheory.2014.07.010-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000913_j.mechmachtheory.2014.07.010-Figure1-1.png",
        "caption": "Fig. 1. Principle scheme of band saw.",
        "texts": [
            "We also can draw the surfaces showing the deformations of themain shaft in twomutually perpendicular planes. These deformations are a function of two parameters, the time t and the coordinate z, which takes into account the length of the shaft. Obtained analytical expressions and diagrams allow an optimal choice of the operating parameters. Thus the normal operation of the band saw machines can be guaranteed. icho_marinoff@abv.bg. 0 2. Expose 2.1. Principle scheme of the band saw machine The scheme of the band sawmachine is shown in Fig. 1 [1\u20136]. The following symbols are defined: 1, 2, 5, 6\u2014 belt pulleys, E\u2014 electric motor, 3 and 4 \u2014 leading wheels, A \u2014 band-saw blade, and 7 and 8 \u2014 chain-wheels. 2.2. Transmission dynamic model The dynamic model is shown in Fig. 2 [7,8]. This model is used to solve the problems. We choose the following coordinate systems [7,8]: Fixed coordinate system O3xyz, moving coordinate system O3x1y1z1, which moves along with the driving wheel. In the initial moment (\u03c6 = 0) the axes of the two coordinate systems coincide"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002864_0954405416661003-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002864_0954405416661003-Figure3-1.png",
        "caption": "Figure 3. Cross sections of the spiroid pinion tooth surfaces for (a) Profile I and (b) Profile II.",
        "texts": [
            " The modified tooth geometry of gear will be offered as die cavity surface that is manufactured by electrical discharge or CNC machine directly. 4. The geometry of double-crowned gear will be manufactured by precision casting process using such modified geometry of die cavity. Geometry of the imaginary generating fully conjugated spiroid gear In theory, the fully conjugated tooth surfaces of spiroid gear can be obtained by using a spiroid cylindrical pinion as a virtual hob to produce the gear as shown in Figure 3. The equations of the tooth space surfaces of left-hand involute spiroid pinion can be described directly. The tooth surfaces I and II of the involute pinion generate the concave and convex tooth surfaces of the gear, respectively (Figure 3). The tooth surfaces I and II of the pinion are represented by the following equations r1I, II = 7 rb1 sin (u01 + u1) u1 cos (lb1) cos (u01 + u1)\u00bd rb1 cos (u01 + u1)+ u1 cos (lb1) sin (u01 + u1)\u00bd 7 u1 sin (lb1) p1u1\u00bd 2 4 3 5 \u00f06\u00de The unit normals to the tooth surfaces I and II are determined in coordinate system S1 by the following equation (7) n1I, II= 6 sin (lb1) cos (u01 + u1) sin (lb1) sin (u01 + u1) 6 cos (lb1) 2 4 3 5 \u00f07\u00de where the upper and lower signs correspond to I and II tooth surfaces for equations (6) and (7)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000866_iros.2014.6943044-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000866_iros.2014.6943044-Figure1-1.png",
        "caption": "Fig. 1. A Mobile Robot Platform using Double-Wheel Caster Units with Servo Brakes",
        "texts": [
            " One of the most important features is casters flexibility; they can move in all directions. Second, when M. Saida, Y. Hirata and K. Kosuge are with Department of Bioengineering and Robotics, Tohoku University, 6-6-01, Aoba, Aramaki, Aoba-ku, Sendai 980-8579, JAPAN {saida, hirata, kosuge}@irs.mech.tohoku.ac.jp transporting an object on a moving base equipped with casters, we can select the number and positions of casters most appropriate to the weight or shape of the object. The number and positions of the caster units are also userselectable in this way, as shown in Fig. 1(c). On the other hand, when a user-controlled mobile robot platform is constructed using casters, the geometrical relationships among the caster units, i.e., the poses of the casters, are essential for odometry and control [5]. Generally, the geometrical relationships of the robot are constant, therefore a user can input them into the controller in advance. In our system, however, after attaching the caster units to the mobile base, the user is required to input the geometrical relationships into the controller before deploying the robot",
            " Moreover, the pose of caster units can be easily and appropriately estimated from a simple movement pattern that includes sufficient state information. The simple moving pattern proposed here enables users to obtain the attachment positions and orientations among the caster units. Incidentally, since the proposed estimation method utilizes only the velocity information of each caster unit, it can be applied not only to our passive systems but also to activeactuated systems. The mobile robot platform using two double-wheel caster units with servo brakes is presented in Fig.1(a). The caster unit is installed with three encoders attached to two wheel shafts and a pivot shaft, as shown in Fig. 1(b). These encoders provide velocity information of the caster units. We first assume a rigid moving base with firmly attached caster units. The coordinate systems of j-caster units are illustrated in Fig. 2. One of the caster units, labeled the k-th caster unit, is selected as a representative unit. We consider the relationship between the i-th caster unit, and the representative, k-th caster unit. In Fig. 2, \u03a3i is the i-th coordinate system (where i = 1, . . . , j) originated at the pivot shaft of the i-th caster unit, and \u03a3k (k 6= i) is the coordinate system of the k-th caster unit"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000047_ijhvs.2020.104400-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000047_ijhvs.2020.104400-Figure1-1.png",
        "caption": "Figure 1 The contact points and unit vectors of the wheelset/rail system (rear view)",
        "texts": [
            " The kinematics analysis includes six subsections: reference frames, generalised speeds, constraint equations, partial angular velocities and partial velocities, linearisation, and angular and linear accelerations. Similarly, the kinetics analysis consists of generalised active forces and generalised inertia forces. The obtained equations of motion based on Kane\u2019s method are compared with those in the literature. At last, the variable parameter of the wheelset system at critical speeds will be investigated. The left and right wheels with conicity angles L\u03b4 and R\u03b4 (Figure 1) are rigidly mounted on a rigid common axle to form a wheelset and these angles are measured between the tangential direction of each contact point and the axial direction of the wheelset. Research that focused on the lateral motion of the wheelset system typically assumes that the wheelset is moving along straight rails at a constant speed V . The wheelset is also assumed to be kept in close contact with the rails during motion, and the contact points between the wheels and the rails are denoted as P and Q , respectively. Two orthogonal unit vector systems, i.e., ( 1, 2, 3)i i =n and ( 1, 2, 3)i i =a , constitute an inertial reference frame N and an intermediate reference frame A (Figure 2), respectively. The mutually perpendicular unit vector system, ( 1, 2, 3)i i =b , shown in Figure 1 denotes a reference frame B which has its origin at the mass centre G and each aligned with a principal axis of the wheelset. The reference frame A is formed by rotating ( 1, 2, 3)i i =n anti-clockwise through a yaw angle \u03c8 about unit vector 3n ; likewise, the reference frame B is formed by rotating ( 1, 2, 3)i i =a anticlockwise through a roll angle \u03d5 about 1a . Table 1 presents the direction cosines matrices between these two sets of unit vector systems: ( 1, 2, 3)i i =n and ( 1, 2, 3)i i =a , together with ( 1, 2, 3)i i =a and ( 1, 2, 3)i i =b , respectively",
            " The relationships between ( 1, 2, 3)i i =n and ( 1, 2, 3)i i =b now are 1 1 2 3cos sin cos sin sin\u03c8 \u03c8 \u03d5 \u03c8 \u03d5= \u2212 +n b b b , (1) 2 1 2 3sin cos cos cos sin\u03c8 \u03c8 \u03d5 \u03c8 \u03d5= + \u2212n b b b , (2) 3 2 3sin cos\u03d5 \u03d5= +n b b (3) Note that ( 1, 2, 3)i i =l and ( 1, 2, 3)i i =r represent the orthogonal unit vectors at P and Q , respectively and Table 2 depicts the direction cosines matrices between ( 1, 2, 3)i i =l and ( 1, 2, 3)i i =b along with ( 1, 2, 3)i i =r and ( 1, 2, 3)i i =b , respectively. Assume that another reference frame D is attached to the wheelset (Figure 1). The angle \u03c6 denotes the forward spinning angle of the wheel and the perturbation angular displacement from a nominal value about 2b is neglected. Based on the concept of simple angular velocity (Roithmayr and Hodges, 2016), the relative angular velocities between two reference frames of N , A , or A , B , or B , D are 3 1 2, , ,N A A B B D\u03c8 \u03d5 \u03c6= = =\u03c9 n \u03c9 a \u03c9 b (4) where the superscripts on angular velocities stand for the rigid body or the reference frames. Now, if the angular velocity of D with respect to N is taken as 1 1 2 2 3 3 N D u u u= + +\u03c9 b b b (5) and the following auxiliary equation is used, then the relationship among generalised speeds ( 1, 2, 3)iu i = and time derivatives of angles \u03c8 , \u03d5 , and \u03c6 are 3sec u\u03c8 \u03d5= (7) 1u\u03d5 = (8) 2 3tanu u\u03c6 \u03d5= \u2212 (9) Likewise, the velocity of mass centre G of the wheelset is now taken as 4 1 5 2 6 3 ,N G u u u= + +v n n n , (10) where the relationships between the generalised speeds ( 4, 5, 6)iu i = and the time derivatives of the coordinates x , y , and z of G are 4 5 6, , ",
            " After performing some vector operation, the velocities at contact points P and Q can be expressed as 3 5 0 1 1 6 5 2 1 6 5 3 [ ( tan ) sin (cos / )] {[( )sin cos ] sin( ) cos( )( cos sin )} {[( ) cos sin ] cos( ) sin( )( cos sin )} N P L L L L L L L L L L L L L L L r a u u V r r a r u u u V a r u u u V \u03d5 \u03c8 \u03c8 \u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03d5 \u03c8 \u03c8 \u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03d5 \u03c8 \u03c8 = \u2212 + \u2212 \u0394 + + \u2212 + \u2212 \u0394 + + + + + \u2212 + \u2212 \u0394 \u2212 + + \u2212 + \u2212 v l l l (19) 3 5 0 1 1 6 5 2 1 6 5 3 [ ( tan ) sin (cos / )] {[( )sin cos ] sin( ) cos( )( cos sin } { [( )cos sin ] cos( ) sin( )( cos sin } , N Q R R R R R R R R R R R R R R R r a u u V r r a r u u u V a r u u u V \u03d5 \u03c8 \u03c8 \u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03d5 \u03c8 \u03c8 \u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03d5 \u03c8 \u03c8 = \u2212 \u2212 \u2212 \u0394 + + \u2212 + + \u0394 + \u2212 \u2212 + \u2212 \u2212 + \u2212 + \u0394 \u2212 + \u2212 + \u2212 \u2212 v r r r (20) where Lr and Rr are written as (Figure 1) 0 tanL Lr r y \u03b4= + , 0 tan .R Rr r y \u03b4= \u2212 (21) The wheels of the wheelset are assumed to be always in contact with the rails, therefore, the velocity components at contact points in their normal directions are zero. That means 3 0N P \u22c5 =v l (22) 3 0.N Q \u22c5 =v r (23) Substituting equations (19) and (20) into equations (22) and (23) yields 1 6 5[( ) cos sin ] cos( ) sin( )( cos sin )L L L L L La r u u u V\u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03d5 \u03c8 \u03c8\u2212 \u0394 \u2212 + + = + \u2212 (24) 1 6 5[( ) cos sin ] cos( ) sin( )( cos sin )R R R R R Ra r u u u V\u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03d5 \u03c8 \u03c8\u2212 + \u0394 \u2212 + \u2212 = \u2212 \u2212 \u2212 (25) Solving equations (24) and (25) simultaneously arrives at 1 1 5( cos sin )u u V\u03c8 \u03c8\u0394 = \u2212 \u0394 (26) 6 6 5( cos sin ),u u V\u03c8 \u03c8\u0394 = \u2212 \u0394 (27) where ,\u0394 1,\u0394 and 6\u0394 are [( ) cos sin ]cos( ) [( ) cos sin ]cos( )L L L L R R R R R La r a r\u03b4 \u03b4 \u03b4 \u03d5 \u03b4 \u03b4 \u03b4 \u03d5\u0394 = \u2212 \u0394 \u2212 \u2212 + + \u0394 \u2212 + (28) 1 sin( )L R\u03b4 \u03b4\u0394 = + (29) 6 [( )cos sin ]sin( ) [( )cos sin ]sin( )"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003157_etfa.2016.7733544-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003157_etfa.2016.7733544-Figure1-1.png",
        "caption": "Fig 1. Application of the proposed system for screwing task",
        "texts": [
            " The robot programming for a change in environment needs consideration of several aspects including object position and orientation, collision avoidance, task related optimization procedures etc. Moreover, the desired task specific accuracy and precision should not be compromised. To address these problems, a new system for a moveable robot integrating a vision system is proposed in this paper. The system is targeted towards use in manufacturing operations with the robot adapting to new scenario for a semi structured environment. A proposed application of the system for screwing task is shown in Figure 1. In this case, the system can be moved arbitrarily to a new workstation by the operator and the robot would perform the task of screwing using the inputs from the vision system. The paper is structured as follows: Section II discusses the current systems; section III presents the implementation of adaptation to environment by the robot, section IV discusses the object recognition; section V discusses the application of the proposed system to assembly task; section 978-1-5090-1314-2/16/$31.00 \u00a92016 IEEE VI discusses the implementation of the system for accurate object detection and finally section VII gives the conclusion and future work"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001797_humanoids.2015.7363511-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001797_humanoids.2015.7363511-Figure5-1.png",
        "caption": "Fig. 5. An illustration of the swing phase.",
        "texts": [
            " When the swing foot disengages contact with the ground, the controller enters the lifting phase, in which the robot lifts the swing leg toward some target position, which is determined from the target stepping point. Researchers have developed a number of methods for determining a stepping point such as the capture point [13] and GFPE [11]. While any reasonable method should work fine, we adopted the method of [12] that determines the stepping point from the desired CoP (See Sec. IV). The target point sproj for the swing foot is determined as shown in Fig. 5. Given the target stepping point sdes, the vector from CoM to sproj is aligned to sdes and the distance of sproj from CoM is set by some desired length L. The desired length L is computed as the length from CoM to the support foot, and then down-scaled to its 80% so that the robot can raise the swing foot. Using sproj gives an advantage that the controller does not need to consider the trajectory of the swing foot while generating reasonably natural stepping motion. Hence, the desired swing foot configuration T d \u2208 SE(3) is determined such that the position equals sproj and the orientation is set to identity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001862_065018-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001862_065018-Figure1-1.png",
        "caption": "Figure 1. Geometry of a biased ball rolling about a horizontal axis at angular velocity \u03c9, and precessing at angular velocity \u03a9 about a vertical axis located outside the ball. The centre of mass, G, is displaced horizontally by a distance d from the geometric centre of the ball, and the normal reaction force, N, is displaced towards the front of the ball by a distance D. F and F\u0302 are friction forces acting tangentially and perpendicular to the path of the ball.",
        "texts": [
            " The trajectory of a biased ball has been described previously in relation to the gentle arc followed by a ball in lawn bowls [9]. A brief summary of that model is reproduced here in part as it relates to the present experiment. The model is a simplified version of the more rigorous treatment of the problem given by Brearley and Bolt [10] and by Brearley [11], but is sufficiently accurate for the present purposes and is consistent with the experimental results. Suppose that a biased ball of mass M and radius R is rolling on a horizontal surface with an initial velocity v0 along the x axis, as shown in figure 1. The ball centre of mass, G, is offset horizontally by a small distance d from the geometric centre of the ball. As a result of precession, the ball follows a curved path rather than a straight line path. The vertical precession axis does not pass through G, as it does with a spinning top, but lies well outside the ball. G itself does not rotate about a vertical axis through the geometric centre of the ball. Rather, the whole ball rotates slowly about the remote precession axis and G remains at an approximately fixed distance from the precession axis",
            " F acts in a direction opposite the direction of motion. Consequently, the velocity, v of the centre of mass is given by v v gt. 10 r ( )m= - The ball comes to rest after a time T v g0 r( )m= . Since the ball follows a curved path, it is also subject to a centripetal force F Mv rG 2=^ where rG is the radius of curvature of the centre of mass. Given that r dG  in the present experiment, we will assume that r rG = in the following discussion, where r is the radius of curvature of the contact point, as shown in figure 1(a). The centripetal force arises as a result of static friction provided that v is not too large, otherwise the ball will commence to slide. Since F\u0302 acts in a direction perpendicular to the path of the ball, it has no effect on the magnitude of v. The ball rotates about the horizontal axis at angular velocity \u03c9. The force component F acts to decrease v linearly with time, but it also exerts a torque F R that acts to increase \u03c9. Given that the ball rolls with v Rw= , \u03c9 must actually decrease linearly with time, indicating that some other effect must be acting to decrease \u03c9. A simple and well-known explanation [1, 2] can be found if it is assumed that the normal reaction force, N, acts a distance D ahead of the point of support, as shown in figure 1(b). In that case, F R MgD I td dcm w- = , where Icm is the moment of inertia of the ball about the rolling axis. Since F Mg M v td drm= = - and v Rw = , it is easy to show that D R I MR 1 , 2r cm 2 ( )\u239c \u239f\u239b \u239d \u239e \u23a0m= + and hence D R 1.4 rm= for a solid, spherical ball with I MR0.4cm 2= . The torque due to the finite value of D is a factor D R 1.4r( )m = times larger than the torque due to the friction force. The offset distance D is related to the extent of deformation of the ball and the surface on which it rolls, thereby accounting for the fact that rm is relatively large for soft balls or soft surfaces, and relatively small for a hard ball rolling on a smooth, hard surface",
            " The deformation itself may be approximately symmetrical in that region, but if the ball is rolling into the surface at the front edge and out of the surface at the rear, then the normal reaction force at the front edge will be larger than that at the rear since the front edge is being compressed and the rear edge is expanding elastically [12]. The component of the angular momentum parallel to the horizontal rolling axis is Icmw, and its rate of change is equal to the torque about the horizontal axis parallel to v. The ball therefore precesses about a remote vertical axis at angular velocity \u03a9. As shown in figure 1(c), a tangent to the path makes an angle f with the x axis, and td dfW = . The ball will roll without slipping provided that v R rw= = W. The ball may also rotate about the horizontal axis parallel to v, especially if d is large, but it is found experimentally that the rate of rotation is relatively small when d is small. Consequently, we will assume for simplicity that G remains on the horizontal rolling axis, at least during the period of experimental interest before the ball comes to rest. The torque about the centre of mass in figure 1(a) is Mgd Mr R I , 32 cm ( )w- W = W where IcmwW is the rate of change of the component of the angular momentum parallel to the rolling axis. Since r Rw = W , equation (3) can be expressed in the form t MgdR I vd d , 40( ) ( )fW = = where I I MR0 cm 2= + is the moment of inertia of the ball about an axis through one edge. The precession frequency therefore increases with time as v and \u03c9 decrease. In theory, W  \u00a5 as v 0 . The predicted behaviour of the ball in this limit is not realistic, and is not valid since the centripetal force in equation (3) vanishes when r = d or when the precession axis passes through G"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001338_028004-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001338_028004-Figure3-1.png",
        "caption": "Figure 3. Schematic diagram showing a cricket ball viewed from above and illustrating the principle of conventional swing with a new ball. In this case the seam is pointed towards fine leg and back-spin is applied to keep the seam vertical. The velocity and angular velocity vectors, together with the drag force and lift (Magnus) force vectors are shown. The wake is deflected in the positive y direction and the ball swings in the negative y direction, that is in the direction to which the seam is pointing, not as a result of the spin itself but as a result of the behaviour of the boundary layer and, in particular, the separation points. The final boundary layer separation points are marked with filled red circles but the actual positions of all separation points are intended to be indicative only (see text).",
        "texts": [
            " (1) D D D In this equation \u03c1 is the density of air, A is the cross-sectional area of the ball, CD is a dimensionless number known as the drag coefficient, V\u20d7 is the ball\u02bcs velocity relative to the air and of magnitude = \u20d7V V| |, v\u20d7 is the ball\u02bcs velocity relative to the Earth\u02bcs surface and W\u20d7 is the wind velocity, also measured relative to the Earth\u02bcs surface. Note that in the above equation \u20d7 = \u20d7 \u2212 \u20d7V v W( ), and equation (1) is sometimes regarded as the definition of CD. The wind velocity, W\u20d7, will be taken to be zero throughout the present paper. In the range of velocities usually encountered in ballgames the drag coefficient for smooth spheres is approximately constant, with a value of about 0.45, and the critical Reynolds number at which the value ofCD abruptly drops by a factor of about 5 is generally not exceeded (see figure 3 of RR2013).7 However, sports-balls are generally not smooth and the effect of the non-smooth surface is to reduce the critical Reynolds number to a value which may certainly be exceeded under certain circumstances, although the subcritical value of CD remains at \u223c0.45. For the purposes of 6 A complete list of symbols used in this paper is given in the appendix. Note, in particular, the adopted values for the constants: cricket ball diameter = \u00d7 \u2212D 7.20 10 m2 and mass = \u00d7 \u2212m 1.59 10 kg1 , air density \u03c1 = \u22121",
            " The adopted coordinate system is also shown, with the x-axis directed along the pitch, the y-axis perpendicular to the pitch and the z-axis vertically upwards. For the purpose of this study the positions of 3rd man and fine leg are important, since these are the approximate directions to which the seam of the ball is usually pointed in swing bowling. The trajectory shown is that of an \u2018out-swinger\u2019 to the right handed batsman, for which the seam is pointed in the general direction of 3rd man, assuming that conventional swing occurs. Figure 3 shows a \u2018new ball\u2019, characterized by both halves of the ball being \u2018smooth\u20199. At relatively low ball speeds (for fast bowlers), conventional swing may occur. The seam is pointed at an angle of about 20\u00b0 to the line of the pitch, in this case being pointed in the general direction of the field position of \u2018fine leg\u2019. Back-spin is applied along the line of the seam to keep it vertical and to promote stability. For the coordinate system shown in figures 2 and 3, on the positive y side of the ball (i",
            " The spin of the ball, aside from promoting stability, has the not necessarily intended effect of causing a lift or Magnus force to be generated and, if the ball is moving horizontally, this force is in the vertical direction. At later stages in the flight however, as will be discussed in section 5 below, when the ball begins to fall the Magnus force develops a side-ways component, also in the direction to which the seam is pointing, thus aiding conventional swing. Figure 4 is a similar diagram to figure 3, but for a higher ball velocity. As the ball velocity increases, on the negative y side of the ball (i.e. the lower side in the diagram), turbulence occurs in the boundary layer itself, resulting in turbulent separation occurring, without passing through the phases of laminar separation and turbulent reattachment characteristic of the lower ball speeds. The separation point of the boundary layer is thus further forward than for the low velocity case. In addition, the boundary layer on the positive y (upper) side of the ball can also undergo transition from laminar to turbulent flow, but with the processes of laminar separation, turbulent reattachment and finally turbulent separation occurring in sequence. When this occurs the separation point of the boundary layer from the ball on the positive y side is delayed and moves backwards towards the rear of the ball. The result is that the wake is now deflected in the negative y direction and the ball swings in the positive y direction. This is known as reverse swing in which the ball moves in the opposite direction to which the seam is pointing. Note that the spin of the ball has not changed from that in figure 3 and hence the direction of the Magnus force has not changed. Thus, when the ball begins to fall, the Magnus force opposes the movement of reverse swing. It should be noted that since the occurrence of conventional or reverse swing depends on, amongst other things, the ball velocity, then there may be a range of velocities for which there is no swing possible. If this range were to be small, it would appear that the nature of the swing which actually occurs would be very difficult to predict or control"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002425_s11771-014-2021-5-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002425_s11771-014-2021-5-Figure3-1.png",
        "caption": "Fig. 3 Generating motions and applied coordinate systems for cosine gear",
        "texts": [
            " (2014) 21: 933\u2212941 935 The design and generation of new tooth profiles require not only high-quality equipment and special tools for manufacturing of such gear drives, but the development of the existing machine-tool settings. Such settings are standardized, but have to be controlled by CNC technology for each case of design (depending on geometric parameters of the new tooth profiles and generating tools) to guarantee the required quality of the profiles. Rather than studying the macro continuous conjugate surfaces, we propose a method to discuss the micro discrete conjugate point in order to manufacture the cosine tooth profile using the standard involute gear slotting cutter. Figure 3 illustrates the relationship between rack cutter and the cosine gear of the gear generation mechanism. Henceforth, we will consider three coordinate systems: S2 that is rigidly connected to the generating surface (too1 surface)  2, S1 that is rigidly connected to the generated surface (gear surface) 1, and  that is the fixed coordinate system where the meshing of 1 and 2 is considered. The derivation of the kinematic control model of generation of the cosine tooth profiles using the slotting method is based on the following procedure: Step 1: Build the digital model of the cosine tooth surface represented by discrete points",
            " (5) can be rewritten as 1 1 1 1 1 1 1 1 1 1 cos sin sin cos 1 1 o i i i i i o i i i i i x x y y x y                        (6) Supposing that k1i is the slope of the cubic spline curve, which is reconstructed by the discrete points using the curve fitting method, the tangent vector in the coordinate system  (O, x, y) of an arbitrary point on the cosine tooth profile, after which rotating an angle \u03c61i clockwise can be represented as 1 1 1 1 1 01 1 1 1 1 1 cos sin sin cos 1 11 o xi i i i o yi i i i i k k k k k                                   M (7) 3.2.1 Mathematical model of gear slotting cutter As shown in Fig. 5, parametric equation of tooth profile of the involute gear slotting cutter in the coordinate system  2(O2, x2, y2) can be expressed as 2 b 2 b (cos sin ) (sin cos ) x r y r            (8) where rb denotes the radius of the base circle,  represents the generating angle of the involute tooth profile, \u03b1k is the pressure angle of point K, \u03b4 is the parametric variable, which can be then expressed as k    As shown in Fig. 3, the equation of tooth profile of the involute gear slotting cutter in the fixed coordinate system 2(O2, x, y) is represented by 2 2 2 02 1 2 2 2 2( ) ( ) 1 1 o o i i x x y y                    M A A (9) where M02 represents the coordinate transformation matrix from coordinate system 2 to , a is the initial distance between the gear blank and gear slotting cutter, 1 2( )iA and 2 2( )iA denote the rotation and translation matrices from coordinate system 2 to \u03a3\u2032, respectively. Moreover, the following relationship exists: 02 1 0 0 0 1 0 0 1 a           M 2 2 1 2 2 2 cos sin 0 ( ) sin cos 0 0 0 1 i i i i i                A  2 2 2 2( ) 1i i ix y   A Hence, Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000859_1.4027130-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000859_1.4027130-Figure7-1.png",
        "caption": "Fig. 7 Measured transmission error spectra of single-row trochoidal gears (Nr 5 60 rpm) Fig. 8 MBA model of a single-row trochoidal gear (Type A)",
        "texts": [
            "asme.org/pdfaccess.ashx?url=/data/journals/jotre9/930316/ on 04/13/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use increased as the rotational speed of the roller gear increased. However, the increasing rate of the measured TEP-P of the type B gear was due to an increase of the rotational speed that was less than that of the type A gear. 3.3 Transmission Error Spectra. The typical measured transmission error spectra of single-row trochoidal gears (types A and B gears) are shown in Fig. 7. In this figure, mark \u2018\u00fe\u2019 shows the peaks with the integer multiple of the rotational frequency of the roller gear nfr (n\u00bc 1,2,3,\u2026) and mark \u2018\u2022\u2019 shows the peaks with the integer multiple of the rotational frequency of the cam gear nfc. Note that nzrfr, nzrfr 6 fr, and nzrfr 6 fc (which are relative to the gear mesh frequency nzrfr) are the multiples of fr or fc. It is clear from Fig. 7 that, for both types A and B gears, the amplitudes of the peaks with fc and fr are relatively large. The peak with zrfr appeared in the spectrum of the type A gear, while it did not clearly appear in the spectrum of the type B gear. To analyze the effects of the contact ratio e on the transmission errors of the single-row trochoidal gears (types A and B gears), the transmission errors were calculated using multibody analysis (MBA) software. 4.1 MBA Model. A model used in the MBA for single-row trochoidal gears is shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001745_omae2015-41955-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001745_omae2015-41955-Figure2-1.png",
        "caption": "Figure 2. ROTOR OF A HYSTERESIS IPM MOTOR.",
        "texts": [
            " Based on the simulation results, a hysteresis IPM motor drive can be a possible replacement of the standard induction motor drive which can improve the reliability and the performance of standard ESPs. HYSTERESIS IPM MOTOR A hysteresis IPM motor\u2019s rotor has a cylindrical ring made of composite material like 17% or 36% cobalt steel alloy, special Al-Ni-Co, Vicalloy, etc. with high degree of hysteresis energy per unit volume [6, 11]. The rare earth permanent magnets are buried inside the hysteresis ring and the ring is supported by a sleeve made of non-magnetic materials like aluminum which forces the flux to flow circumferentially inside the rotor ring [6, 11]. Fig. 2 illustrates the rotor of a hysteresis IPM motor. The cross section of a hysteresis IPM motor depicting the position and orientation of permanent magnets is shown in Fig. 3. The inclusion of permanent magnets creates rotor saliency without changing the length of the physical airgap and provides an additional permanent source of excitation in the rotor. Magnetic hysteresis refers to the dependency of a material\u2019s magnetization on the past states of the material, as well as its current state. Fig. 4 shows the major hysteresis loop of a hysteresis ring made of 36% Cobalt-Steel alloy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002944_icma.2016.7558946-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002944_icma.2016.7558946-Figure4-1.png",
        "caption": "Fig. 4. Loading process. The nodes are rigidly fixed on one side of the vascular when axial loading (left), and the forces are equally exerted on each node on other side. There are no nodes fixed when radial loading (right), the forces are exerted on each node.",
        "texts": [
            " The biomechanics properties of axial tension and radial expansion of vascular model is analyzed, and the experimental results are compared with the results obtained by Georgia Institute of Technology. The mechanical properties of biological soft tissues are characterized by the viscoelasticity, anisotropy, stress relaxation and creep [9], [10]. When the soft tissues suffer cyclic load, the unloading curve normally falls below the loading curve to form a hysteresis loop. The vascular model\u2019s parameters is empirically adjusted, then two experiments is performed to test the axial and radial mechanics of the model. The loading process is visualized in Fig. 4. The forces during loading and unloading processes vary at a same constant speed. As shown in Fig. 5, the hysteresis is very obvious, but the errors between the results of the BVS and the vascular model are large. Due to this, the parameters of the model need to be optimized. In general, the vascular model\u2019s parameters is empirically adjusted, but it is inefficiency and difficult to achieve the desired effect, and inappropriate parameters can cause severe shock in the model. A better way is to use machine learning algorithm to learn the parameters that can make the model stable, and then to learn the errors between virtual vascular biomechanics property and actual vascular biomechanics property to find the optimal parameters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.11-1.png",
        "caption": "Fig. 10.11 Configuration of (a) isometric and (b) top views of face milling",
        "texts": [
            " Meanwhile, the material sample mounted on a movable stage moves toward the milling head with a feed rate f (millimeter per minute), and the feed distance per single tooth cut (Lfeed_tooth) is f/N/ n (or the feed distance per single revolution Lfeed \u00bc f/N). 278 10 Design for Manufacturing The cutting time Tcut is calculated by Eq. 10.2, whose length of initial offset position (LA) can be further estimated by this simple proof provided below1: LA \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi t D t\u00f0 \u00de p \u00f010:6\u00de where t is the width/depth of cut. On the other hand, the MRR with LA L is MRR LWB Tcut \u00bc fWB \u00bc kv \u03c0D WBLfeed \u00bc WBNnLfeed tooth \u00f010:7\u00de For face milling (Fig. 10.11a), the rotational axis of a milling head with a diameter D is normal to the machined surface. With the same definitions of most process parameters, the machine speed and feed rate are computed as the previous section. The cutting time Tcut should consider both the offset at the beginning (LA) and ending (LO) of the process: Tcut \u00bc LA \u00fe L\u00fe LE f \u00f010:8\u00de where L is length of cut on the sample. In more detailed considerations, the actual values of LA and LE are determined by different situations of the cutting width W. For example (Fig. 10.11b): \u2022 If the cutter is not fully engaged and W is larger than D/2, then LA \u00bc LE \u00bc D/2. \u2022 If the cutter is not fully engaged and W is smaller than D/2, then LA \u00bc LE \u00bcffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi W D W\u00f0 \u00dep . 1As shown in the right inset of Fig. 10.11, since the lengths OA \u00bc OB \u00bc OC \u00bc D/2 (radius of mill head), angle BAO \u00bc angle ABO, and angle OBC \u00bc angle OCB. Considering triangle ABC, 2 (\u03b8 + \u03a6) \u00bc 180 , and thus angle ABC \u00bc 90 . Considering triangle ABX, tan\u03a6 \u00bc LA/(D t). Considering triangle BCX, angle XBC \u00bc \u03a6, and thus tan\u03a6 \u00bc t/LA. Therefore, tan\u03a6 \u00bc LA/(D t) \u00bc t/LA. 10.2 Machining Process Design 279 The MRR can then be approximated as Eq. 10.7, ignoring the offsets of engagement (LA) and over-traveling (LE). One major concern of machining processes is the cost of manufacturing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure10-1.png",
        "caption": "Fig. 10 Mathematical model of the lower and upper leg; a forces are indicated and local coordinate systems of the thigh and shank; b Dimensional parameters and muscle insertion and slide points are presented",
        "texts": [
            " In equations represented by I\u03012 Insertion1b Relative coordinates of knee flexor muscles\u2019 insertion point into femur. In equations represented by I\u03011b Insertion2b Relative coordinates of knee flexor muscles\u2019 insertion point into tibia. In equations represented by I\u03012b animation_time Animation time expressed in seconds. Based on this value, and the knee motion range, the time step between view refreshing is calculated kilograms and the force in Newtons. The visualization of the model is provided in Fig. 10; the forces acting on the model are shown on the left side of the figure, and the dimensions and locations of characteristic points are presented on the right side. Muscle force is divided into the extensor muscle force, Fme, and flexor muscle group, Fm f . It is assumed in the model that muscles can only produce force while contracting. The results of the simulation are the hip and knee joint torques required for keeping the leg system in equilibrium; muscle forces where a positive value means that muscles are contracting; shear and normal forces, and bending moment in the knee as estimated by torque actuated and muscle actuated knee model; and the total knee force, with division on tangential and normal components",
            " The force diagram for the model with muscles is illustrated in Fig. 12. In the presented model, each muscle group can be regarded as a two-segment rope, which is represented in the Fig. 12 as a dotted line. If the muscle segments were virtually cut, the internal muscle forces can be applied to wrapping points and muscle insertion points. Point K then represents the knee cap, while point K \u2032 the knee pit. Points K and K \u2032 are muscle wrapping points and are considered to be rigidly attached to the upper leg at its end. Distances d and d1, depicted in Fig. 10b, represent distances between points P2 and K , and P2 and K \u2032 respectively. The moment equilibrium equation for the lower leg segment can be derived as follows. Mk \u2212 Fme Kx \u2212I2x|I2 K | ( I2y \u2212 P2y ) + Fme Ky\u2212I2y |I2 K | (I2x \u2212 P2x ) = 0, Mk \u2265 0 Mk \u2212 Fm f K \u2032 x \u2212I2bx |I2b K \u2032| ( I2by \u2212 P2y ) + Fm f K \u2032 y\u2212I2by |I2b K \u2032| (I2bx \u2212 P2x ) = 0, Mk < 0 (20) Due to the assumption of no antagonistic muscle effect, Eq. 20 is partitioned depending on the sign of the moment, Mk , which can be computed from Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.4-1.png",
        "caption": "Fig. 3.4 Beam under pure bending in the x\u2013z plane: a moment distribution; b deformed beam. Note that the deformation is exaggerated for better illustration. For the deformations considered in this chapter the following applies: R L",
        "texts": [
            "2, only the Bernoulli beam is considered. Consideration of the shear part takes place in Chap.4. 6The sum of all points with \u03c3 = 0 along the beam axis is called the neutral fiber. 7Gustav Robert Kirchhoff (1824\u20131887), German physicist. 8Eric Reissner (1913\u20131996), German/US engineer. 9Raymond David Mindlin (1906\u20131987), US engineer. 92 3 Euler\u2013Bernoulli Beams and Frames For the derivation of the kinematics relation, a beam with length L is under constant moment loading My(x) = const., meaning under pure bending, is considered, see Fig. 3.4. One can see that both external single moments at the left- and right-hand boundary lead to a positive bending moment distribution My within the beam. The vertical position of a point with respect to the center line of the beam without action of an external load is described through the z-coordinate. The vertical displacement of a point on the center line of the beam, meaning for a point with z = 0, under action of the external load is indicated with uz. The deformed center line is represented by the sum of these points with z = 0 and is referred to as the bending line uz(x)",
            "25) describes the bending line uz(x) as a function of the bending moment and is therefore also referred to as the bending line-moment relation. The product EIy in Eq. (3.25) is also called the bending stiffness. If the result from Eq. (3.25) is used in the relation for the bending stress according to Eq. (3.21), the distribution of stress over the cross section results in: \u03c3x(x, z) = +My(x) Iy z(x). (3.26) 3.2 Derivation of the Governing Differential Equation 99 100 3 Euler\u2013Bernoulli Beams and Frames The plus sign in Eq. (3.26) causes that a positive bending moment (see Fig. 3.4) leads to a tensile stress in the upper beam half (meaning for z > 0). The corresponding equations for a deformation in the x\u2013y plane can be found in [37]. In the case of plane bending with My(x) = const., the bending line can be approximated in each case locally through a circle of curvature, see Fig. 3.10. Therefore, the result for pure bending according to Eq. (3.25) can be transferred to the case of plane bending as: \u2212 EIy d2uz(x) dx2 = My(x). (3.27) Let us note at the end of this section that Hooke\u2019s law in the form of Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002561_mawe.201600333-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002561_mawe.201600333-Figure1-1.png",
        "caption": "Figure 1. Schematic illustration of a master model",
        "texts": [
            " Experiments were also carried out to investigate the flexibility of the system developed. Some good results have been obtained and discussed. The surface morphology of the rapid prototypes after removing process was measured using an optical microscopy, a profilometer and a white-light interferometry. Mechanisms of removing support materials for the general and hollow Rapid prototypes were also investigated and discussed. The schematic illustration of a master model, which was designed by Pro/ENGINEER software is shown, Figure 1. The model was then exported to the fused deposition modelling QuicksliceTM software via the stereolithography format. Once the stereolithography file has been exported to QuicksliceTM, it was then horizontally sliced into many thin sections for fabricating the master model by fused deposition modelling rapid prototyping machine. A master model with the support materials is given, Figure 2. A list of process parameters for fabricating master model is given, Table 1. The fabricated material used in manufacturing the test parts was acrylonitrile butadiene styrene"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003257_iros.2016.7759520-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003257_iros.2016.7759520-Figure5-1.png",
        "caption": "Fig. 5 Gait pattern controller",
        "texts": [
            " suggested the locomotion control system which changes the gaiting patterns of a robot with two or four legs according to its walking speed, stability and adaptability [12], [13]. In the same way, we applied similar control method to ASTERISK as six-legged robot which can easily retain its stability since it can secure more grounding points, which include the case leg failure. Thus, we tried to invent the gaiting strategy which is suitable for traveling on an uneven surface and enabled ASTERISK to continue walking even one of its leg broke down. The controlling system shown in Fig. 4 consists of factors of gaiting pattern controller (shown in Fig. 5) and leg motion controller (shown in Fig. 6). As RG model of CPG, gaiting pattern controller which consists of oscillators arranged for each legs generates basic rhythm for motion of each legs. In addition, it composes of gaiting pattern based on command value of basic gait pattern and sends command signal to leg motion controller. On the other hand, leg motion controller moves joint actuators of legs to put orbital motion given by gait pattern controller into practice. Gaiting pattern controller consists of non-linear oscillators arranged for each legs, furthermore, it composes gaiting pattern based on command value of basic gaiting pattern and sends command signal to leg motion controller as described earlier"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003509_jsee.2016.05.15-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003509_jsee.2016.05.15-Figure1-1.png",
        "caption": "Fig. 1 RCS mechanism",
        "texts": [
            " The rest of this paper is organized as follows. In Section 2, nonlinear equations of dual-control missiles are deduced. Then the structure of the proposed control scheme and the design of the attack angle track controller are presented in Section 3. In Section 4, the stability analysis of the dual-control system error dynamic is given. To verify the effectiveness of the proposed control law, simulation results are provided in Section 5. Finally, conclusions are drawn in Section 6. A missile with RCS is described in Fig. 1. The RCS is located ahead of the center of gravity (c.g.). Ox1y1 is the body coordinate system. lR represents the distance between c.g. and RCS location. The RCS is pulsewidth modulated thrusters placed on Oy1 and \u2212Oy1. Missile airframe dynamics in pitch plane is given as \u03b1\u0307 = \u03c9 + mg cos \u03b8 mv \u2212 a1\u03b1 \u2212 a2\u03b4z \u2212 a\u2032 2uR (1a) \u03c9\u0307 = a3\u03c9 + a4\u03b1 + a5\u03b4z + a\u2032 5uR (1b) In above equations, the symbols a1, a2, a3, a \u2032 3, a4, a5, a \u2032 5 are defined as a1 = qSc\u03b1 mv , a2 = qSc\u03b4z mv , a\u2032 2 = cos\u03b1 mv , a3 = qSlm\u03c9 Jz , a4 = qSlm\u03b1 Jz , a5 = qSlm\u03b4z Jz , a\u2032 5 = lR Jz , m is the missile mass, v is the missile velocity, \u03b4z and uR are tail deflection angle and normalized attitude thruster force, q is the dynamic pressure, Jz is the moment of inertia of the missile, S and l are reference area and length, c\u03b1 and c\u03b4z are differential coefficients of lift, m\u03c9, m\u03b1 and m\u03b4z are differential coefficients of pitch moment, \u03b1, \u03b8 and \u03c9 represent the angle of attack, flight-path angle and pitch rate, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001742_s00542-015-2460-4-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001742_s00542-015-2460-4-Figure6-1.png",
        "caption": "Fig. 6 The contact region between the disk and the ramp",
        "texts": [
            " These springs were located at the center of trailing edge (TE), the leading edge (LE), inner edge (IE) and outer edge (OE) of slider. When an external shock was applied, it was transmitted through four bolt points. The large mass method is one of methods to describe the external shock transmission, whereby a large mass was connected to the four bolt points as shown in Fig. 5. And, the external shock is applied to the large mass to investigate the shock response. With the contact model, it is necessary to establish which surfaces are contact and target surfaces. As shown in Fig. 6, both sides of the ramp were considered to be contact surfaces, and both sides of the disk were target surfaces. The augmented Lagrangian formulation includes controls to automatically reduce the penetration, and supports symmetric behavior. By exploiting this symmetric behavior, the contact and target surfaces were constrained from penetrating each other. Is is important to establish accurate contact modeling. Each contact model was performed using a nonlinear analysis, and the contact model was applied to all case, because ramp\u2013disk contact is nonlinear characteristic"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003857_0954406215589843-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003857_0954406215589843-Figure9-1.png",
        "caption": "Figure 9. Example finite element model of a compliant tilting pad design.",
        "texts": [
            " Nevertheless, if the finite element model is linear, its dynamics can also be expressed in terms of its undamped natural mode shapes, and often only a limited number of those are required to properly capture the dynamics of the model. This modal reduction is acceptable if: 1. There exist linear combinations of the selected mode shapes that accurately represent the displacements due to static loads, e.g. the preload force; 2. The frequency of a neglected mode shape is high compared to the frequencies of the external loads; 3. The mode shape does not get excited by the external loads. As an example, Figure 9 shows a simple finite element model of a typical compliant tilting pad design using a flexure for the tilt motion and a leaf spring at UNIV OF CONNECTICUT on June 4, 2015pic.sagepub.comDownloaded from for the radial support. The model has been created with Ansys and consists of quadrilateral shells, two node beams and a special element that connects a node at the pad centre point to the pad nodes, making it follow the average motion of the pad. The edges of the leaf spring, marked with a red circles in the side view, have been clamped",
            " With this relation the pad model equations of motion can be written as MQ \u20acq\u00bd \u00fe CQ _q\u00bd \u00fe KQ q\u00bd \u00bc Fgravity \u00fe Fpreload \u00fe Fair\u00bd \u00f046\u00de Furthermore, using mass-scaling for these eigenvectors and pre-multiplying the equation with QT results in I \u20acq\u00bd \u00feQTCQ _q\u00bd \u00fe ,2 q\u00bd \u00bc QT Fgravity \u00fe Fpreload \u00fe Fair\u00bd \u00f047\u00de where I is the 6 6 unity matrix and , is a diagonal matrix of the six eigen-frequencies belonging to the columns of Q. The matrix C is often not known for modelled systems, however, it is common practice to replace the matrix QTCQ with 2 ,, where represents the fraction of the critical modal damping. This approach is also used for the pad model, applying a small modal damping of at most 2% for 3D simulations. The modal gravity load QT Fgravity is constant during a simulation, so it only has to be determined once for every pad orientation. Also the preload, as shown in Figure 9, acts on the z-translation of one node and therefore selects and scales only one column of the QT matrix. The pressure and friction loads of the air film, on the contrary, are highly dynamic and in general would require for each time step the conversion of the pad node motion to motion of the grid used to calculate the pressure in the air film. Nevertheless, because the pad is very thick compared to the flexure and leaf spring, the first modes of the pad model mainly show deformation of the compliant support structure, while the pad moves like a rigid body"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002029_iemdc.2015.7409101-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002029_iemdc.2015.7409101-Figure5-1.png",
        "caption": "Fig. 5. Configuration of winding I and winding II of the calculated motor on 12-slot 10-pole concentrated winding permanent magnet motor.",
        "texts": [
            " (10) For given operation point each \u03c9, , Vdc is fixed and for the conclusion the amplitude of carrier harmonic current is defined by \u03c6. Carrier harmonic iron loss is independent of phase current, because carrier harmonic current is only influenced from magnetic flux in low torque region. The proposed motor utilizes two sets of three phase windings which are wound around different teeth each other. As an example, concentrated winding PMSM with 12-slot 10-pole is applied, winding I and winding II are alternately set on teeth as shown in Fig. 5. The motor can be driven with both winding I and winding II (Operation mode I: conventional), or with only winding I or II (Operation mode II: proposed). Proposed motor drive system is operated by operation mode I when required shaft torque is middle or higher. The operation mode II is utilized at low torque region. When the motor is drive at operation mode II, six windings and six teeth are not exited. PWM carrier harmonic iron loss does not exist, because armature winding flux is not through non-excitation teeth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001615_1.4031579-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001615_1.4031579-Figure1-1.png",
        "caption": "Fig. 1 Contacting filament seals: \u201c(a) brush seal; (b) leaf seal; (c) finger seal, Proctor et al. [3]; and (d) compliant plate seal, Deo et al. [11]",
        "texts": [
            " Here a compliant barrier is created between the stator and rotor, which generally is in contact with the rotor either circumferentially or axially. During growth or excursions of the rotor, this compliant barrier is able to move with the rotor, thereby ensuring that the leakage remains low and that neither the rotor nor the barrier elements (called filaments) are significantly abraded. The typical seals that fit into this group are brush seals, leaf seals, finger seals [3], Figs. 1(a)\u20131(c) and possibly the compliant plate seal [4], Fig. 1(d). In order for compliant filament seals to be viable technologies in gas turbines, they need to exhibit a low leakage which is maintained throughout their operational life. To achieve this, the seals must demonstrate a low wear rate for all operating conditions they may be subjected to. Wear, especially during incursions, is very damaging to seal performance, as it can cause seals to operate with an increased clearance at other operating conditions. This results in more leakage, which is a performance loss",
            " Currently brush seals are the most widely used filament seal in turbine applications. However due to the somewhat irregular packing of the bristles within the seal and the presence of interbristle contact, they are challenging to analyze, display significant nonlinearity, and show hysteresis [7,8]. The filaments are axially compressed, causing substantial interelement contact forces [9,10] and even interelement wear [11]. The issues are alleviated in leaf seals, which consist of only a single row of sealing filaments (Fig. 1(b)). This has the advantage that no axial contact between seal filaments exists, thereby simplifying the physical interactions within the seal significantly. Hence, the current investigation is focused on these as a step toward general understanding of the stiffness behavior of axially in-contact radially compressed filament seals. An alternative recently invented seal that may have similar attributes to the leaf seal is the compliant plate seal [12]. The seal arrangement in this case contains a central divider (Fig. 1(d)), which directly affects the air flow through the seal, and it is designed to operate as a hydrostatically air-riding seal [4]. Data for experimental test campaigns on leaf seals have been published by Nakane et al. [13] and Jahn et al. [14,15]. In the latter, the first quantitative evidence of aerodynamic forces having a significant impact on seal stiffness was shown. In addition, an empirical model for the mechanical and aerodynamic forces, which are the cause of leaf rotor contact forces, was presented by Jahn et al",
            " Hence by setting relevant targets for the mechanical and aerodynamic constants, the seal mechanical and aerodynamic aspects of seal design can be decoupled. A downside of the empirical model is that it is not possible to experimentally decouple the two aerodynamic forces acting on the seal element and thus no understanding of their relative magnitudes or respective causes can be developed. Hence in order to advance the understanding of these forces and their causes, essential when designing leaf seals to have optimum properties, a numerical approach is required. Out of the filament seals shown in Fig. 1 (brush seal, leaf seal, and finger seal), the leaf seal lends itself best to numerical analysis. This is the case as axial deflections of the filaments are almost eliminated and as contact is restrained to the direction perpendicular to the leaf surface. This allows them to be treated as quasi-2D during mechanical modeling. Similarly, the modeling of the flow through the interleaf gaps lends itself to a number of simplifications. To accurately model the coupled aeromechanical response of the leaf pack to a given flow condition it is required to determine the pressure distribution within the leaf pack and to predict the flow turning that takes place as the flow enters the leaf pack"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.24-1.png",
        "caption": "Fig. 3.24 Relative kinematics of the spatial four-link mechanism",
        "texts": [
            " The assembly of the complete system can be supported with computer aid through suitable algorithms. The global constraint equations can also be represented with use of nonlinear transmission blocks for the individual loops, as well as linear summing points for the coupling equations, well-arranged with help from block diagrams. This common approach, taken from control theory, is exemplified in Sect. 3.5.5, again using the example of the double-wishbone suspension from automotive engineering introduced earlier in this chapter (Fig. 3.24). For the position analysis of the nonlinear transmission motion, e.g. the rotation of the input crank (angle b1) can be chosen as the independent input coordinate q \u00bc b1. The deflection of the right output lever (angle b7) such as the motion of the connecting coupling bar d can be calculated. In particular, one can show that the output variable, which is to be calculated in the first step, b7 \u00bc b7\u00f0b1; geometry of the initial position\u00de; \u00f03:14\u00de can be found explicitly through the solution of the equation Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure10-1.png",
        "caption": "Figure 10 Proposed manipulator (see online version for colours)",
        "texts": [],
        "surrounding_texts": [
            "Figure 15 shows the developed duplex manipulator. Figure 16 shows the locking mechanism. Figure 17 shows the decoupled manipulators. Figure 18 shows the pulling mechanism. Table 1 shows the parameters of developed robot. As shown in Figure 16(a), when the hose expands, it pushes the pin until the pin engages with the hole, which locks the joint. Figure 16(b) shows the resolution of the locking mechanism, and Figure 16(c) shows the pistons to expand the hose. As shown in Figure 18, pulling the wire of the locked manipulator pulls the movable manipulator. Figure 19 shows the camera and four LED lights. We employed a UCAM-DLK130T (Elecom) camera, which is used by the operator to search for survivors. We conducted some experiments to demonstrate the effectiveness of the developed duplex manipulator. First, we confirmed the movement of the head link. Figure 20 shows the experimental result for one manipulator. The head could move up to a maximum of about 30 cm in the vertical direction. The direction of the head link was controlled by pulling the wires. Next, we confirmed the movement of the proposed duplex system. We placed a small white cylinder in a threedimensional environment as a target object, and an operator controlled the duplex manipulator toward the target. In this experiment, the operator could see the manipulator and target directly. Figures 21 to 23 show the experimental results. The proposed manipulator could turn in the three-dimensional environments and reach the target. In particular, it could turn without using the wall as a guide, as shown in Figure 22. Hence, we confirmed that the proposed manipulator can solve the problem described in Section 2. The basic performance of the manipulator is listed in Table 2. Finally, we confirmed the searching capability in rubble. Figure 24 shows the experimental environment. A small white cylinder was set in this rubble as a target, as shown in Figure 25. In this case, the operator did not know the position of the target in advance. A camera was mounted on top of the manipulator, and the operator searched for the target based on images from the camera. The manipulator was inserted from the right side of the environment, and we measured the time required for searching. Figure 26 shows the realised path of the manipulator, and Figure 27 show images from the camera at the positions shown in Figure 28. The average time required for searching was about 3 min. We concluded that the proposed manipulator is useful for search and rescue operations. In these experiments, the environments were artificial and restricted to particular examples. In the future, we intend to evaluate the general performance of our proposed manipulator in real rubble. In this study, we developed a robotised endoscope with the advantages of both robots and industrial endoscopes. We improved our previous manipulator for survivor searches and developed a new duplex manipulator that can proceed in any desired direction through rubble. This duplex manipulator has four important features. First, the duplex mechanism consisting of two semi-circular manipulators with a locking mechanism allows the developed manipulator to proceed in any desired direction. Second, the pulling mechanism works with pushing to greatly reduce deadlock due to friction between the two manipulators. Third, the simple operation repeats only three steps: 1 turn the head link to the desired direction 2 lock all joints of one manipulator 3 move the other manipulator forward along the locked manipulator. Fourth, the low energy consumption since the manipulator can be moved by human power; only the camera and LED light require small amounts of electrical power which can be supplied by small batteries. Thus, the manipulator can be operated even in a massive blackout. To demonstrate the effectiveness of the developed duplex manipulator, we conducted experiments in artificial rubble. The results showed that the duplex manipulator could change its moving direction in three dimensions and was able to reach the target object. We concluded that the developed duplex manipulator retains the advantages of an industrial endoscope while having greater mobility. In future work, we will focus on expanding the searching capability by adding sensors for sound source localisation, heat, etc."
        ]
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure3-1.png",
        "caption": "Figure 3 Photograph of the pinion-type cutter used in the experiment.",
        "texts": [
            "963 mm 2 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/86626/ on 03/02/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use In an actual skiving process, factor such as vibration due to the variation of cutting resistance can influence the accuracy of the skived gear. Therefore, in order to exclude such factor, material that has low cutting resistance such as carbon was employed as the material of the blank. Table 2 shows the pinion-type cutter data and Fig.3 shows a photograph of the cutter. The cutter accuracy was of class JIS AA. To generate internal gear teeth through whole face of the workpiece, a cutter with synchronous rotation must be fed in the direction of the axis of the workpiece. Therefore, spiral cutting marks are formed on the tooth flanks, and its pitch increases in proportion to the feed rate, which deteriorates the gear accuracy as shown in Fig. 4. However, because internal gear and cutter have a high contact ratio, cutter with pitch deviations, directly influences pitch deviations of the skived gear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001842_amm.658.377-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001842_amm.658.377-Figure2-1.png",
        "caption": "Fig. 2 Temperature gradient according Jacq anomaly [6]",
        "texts": [],
        "surrounding_texts": [
            "Rolling linear contact is significantly present in a number of industrial applications, such as rolling bearings, gears, friction wheels, rolling mill rollers. For machine parts located in an enclosure, where premature deterioration is difficult to register, for a continuously working program, under important thermomechanical contact loads, reliability is very important to be at an optimum level. The decisive process in the destruction of such surfaces in rolling linear contact with a small slide, without an imposed lubrication, is global thermomechanical wear [2], an atypical and complex process of different types of wear, with thermal or mechanical causes. In order to obtain for such contacts all their potential lasting in exploitation, it is essential to know the mechanism of destruction, with both mechanical and thermal aspects. It is a complex process with many factors and determinations. This combination of factors has a conjugate result in destruction of the contact layers. The thermomechanical contact fatigue is the major cause of contact surfaces deterioration, with a certain presence in every rolling contact, under thermal tide and mechanical load. Knowing the position of the deterioration origin points is very important in order to take possible preventive measures for the destruction process. The thermomechanical destruction of the contact surface under combined thermal and mechanical loads is separately studied, for each cause. The critical mechanical stresses causing the primary destruction in rolling linear contact, have two different positions, according with the technical literature concerning the rolling linear contact under mechanical load. [1, 7] Critical stresses in rolling contact mechanical fatigue The mechanical contact fatigue is the main deterioration cause for the parts of machine with a dominant contact rolling solicitation in service, under mechanical load (rolling bearings, gears). All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 129.128.216.34, University of Alberta, Edmonton, Canada-26/04/15,13:27:48) There are two kinds of different hypotheses concerning the localization of the critical stresses, presented in [3]. These stresses are localized on the surface or below the contact surface, at different depths. On the contact surface, the maximum normal stress (\u03c30, \u03c3max) is located by McKelvey, Mayer and Neifert, the maximum tangential stress (\u03c445D), with two points of view, by S.V. Pineghin with \u03c445(xy)sb, Foord, Hingley, Cameron and Cioclov with \u03c445D(xy)sa (\u201csa\u201d and \u201csb\u201d indicate the position on the contact ellipse), and the traction normal stress \u03c3yt by Moyar, Morrow and Pineghin. Below the contact surface, the tangential orthogonal maxim stress (\u03c40) can be noticed and was taken into account by Lundberg and Palmgren [4] as the maximum for \u03c4yz, localized under the contact surface, at the depth of z0; the critical tangential stress (\u03c4c) by Ollerton, Morey, Stullen and Cummings is (where kc has values determined by the number of load cycles): \u03c4 = \u03c4 + k \u2219 \u03c3 , with \u03c4yz = \u03c40 (1) \u03c3 = max(\u03c3 , \u03c3 , \u03c3 ) (2) Also, the equivalent critical stress (\u03c3ED) developed by Popinceanu, Diaconescu and Cretu [5] using in their theory an equivalent stress presented by Huber \u2013 Misses \u2013 Heuckey, is given as: \u03c3 ( ) = (\u03c3 \u2212 \u03c3 ) + (\u03c3 \u2212 \u03c3 ) + (\u03c3 \u2212 \u03c3 ) + 6\u03bb \u03c4 \u2212 \u03c4 + 6\u03c4 (3) where \u03bb is the value of the ratio of different fatigue stress limits, from normative and according to the type of variable solicitation. The equivalent critical stress (\u03c3ED) has the prospect to evaluate a complex load situation, as the one which determines the thermomechanical contact fatigue, containing components of both thermal and mechanical origin causes. Effect Jacq in thermal conduction In 1961, J. Jacq, at Heat Transfer Congress in Paris, presented his first experimental test concerning a thermal disturbance in heat thermal conduction [6]. These experimental results show that in metal conductors the thermic field is different than the thermic field given by the Fourier law for a conductivity \u03bb, equal in all the thickness of the metal. The temperature fall is located under the surface of the metal wall at a very small depth on both surfaces of the wall ( on the entry and exist surfaces of the thermal flow ). The hypothesis using as decisive stresses \u03c30, \u03c445D and \u03c3yt, that have the maximum values on the contact surface, gives no explanation for the origin of the destructions under the contact surface, fact demonstrated when the \u03c40, \u03c4c and \u03c3ED(\u03bb) hypothesis are taken into account. The equivalent stress \u03c3ED(\u03bb) has a high enough value on the contact surface, thus giving a global explanation for the destruction origin points, both on and under the contact surface. In Fig. 1 and 2 [6], the notifications have the follow means: tp.cl = wall temperature, following Fourier\u2019s law, tp.r = the real temperature distribution for the wall, taking into account the thermal Jacq effect, \u2206ti = tpr1 \u2013 tpr2, the thermal fall after Jacq effect, \u2206tcond is the thermic fall in conduction transfer (the classically calculated value for the temperature). The thermal gradient is different in these two methodologies. The importance of this thermal effect for the position of primary destruction point is that the change of slope in the temperature variation very near to the surface indicates a variation of thermal stress in that region. Hence, below the surface, between very near levels, different thermal stresses are generated. The material structure, at a certain depth, will be supplementary loaded and will produce a very small crack, which can be considered as an origin of destruction for the thermo-mechanical loaded structure. The range for the temperature difference between very near depths can be illustrated by Fig. 3 [6, A non-stationary regime of thermal tide increases the temperature fall in the Jacq layer, concentrated at a depth of only 50 \u00b5m from the surface (fig 4). For a very high thermal tide on the metal structure, there is possible to obtain thermal falls more than 100 \u00b0C at the depth of 100 \u00b5m [7]. This situation is typical for rolling mills rollers, where the temperature at the active surface could reach around 1000 \u00baC. The deterioration of the rolling mill rollers, illustrated by cracking fields and exfoliation zones on the contact surface, has the origin below the surface due to both unsteady mechanical and thermal stresses. The influence of thermal Jacq effect will intensify this process."
        ]
    },
    {
        "image_filename": "designv11_64_0001149_2014-01-1797-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001149_2014-01-1797-Figure12-1.png",
        "caption": "Figure 12. Architecture of lithium-ion battery",
        "texts": [
            " As described in section 2, only the specifications of the short term rates were changed from the first-generation system. The maximum motor torque available instantaneously at engine start was boosted to 290 Nm from 270 Nm only when starting the engine. This was accomplished by confirming motor durability in relation to the 20 Nm increase in motor torque. The results confirmed are described in section 3 and 4. The short-term maximum output power of the battery at engine start was boosted to 75 kW from 60 kW. The original output level of 60 kW was already high thanks to the lithium-ion battery. Figure 12 shows the battery architecture. This battery has high heat transfer performance due to its laminated cells and can be charged and discharged stably and quickly due to its manganese-based cathode with a spinel structure. For the second generation of the hybrid system, the maximum output power was limited precisely according to the discharge duration in order to boost the short-term maximum output power. Figure 13 shows the architecture of clutch-1. Clutch-1 is a dry clutch that is normally engaged in the EV mode"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002967_aim.2016.7576920-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002967_aim.2016.7576920-Figure2-1.png",
        "caption": "Figure 2. Simulation environment.",
        "texts": [
            " \u03c1 is the coincidence parameter. 1 2,\u03b5 \u03b5 are shape parameters and can be determined after shape recognition. Finally, the object model can be expressed as ( , , )x y z= \u22c5 +x R D T (11) After all the parameters are obtained, object models can be recovered by superquadrics function. The whole model recovery process is shown in Fig. 1. Tactile point clouds are obtained through both simulation system and humanoid robot platform. The simulation environment is built by visual C++ and OpenInventor. As is shown in Fig. 2, it contains a robot arm and a five-fingered robot hand model. When fingers are contacted with object, the location and the normal vector of contact point are displayed and recorded. The humanoid robot, which consist of a pair of humanoid arm and robot hand, is used in the exploration experiments, as is shown in Fig. 3(a). The positional repetitive accuracy for the end of the arm is around 0.2 mm. The five fingered robot hand DLR/HIT Hand II [21] is multisensory and highly integrated, each finger has 3 DOFs and four joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure6.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure6.9-1.png",
        "caption": "Fig. 6.9 Plate problem with four edges fixed",
        "texts": [
            " The solution of this system of equations gives: u2Z = u3Z = \u2212 2(3\u03bd2 + 2\u03bd \u2212 6)a2F 3D(3 + 2\u03bd)(\u22121 + \u03bd) , \u03d52X = \u2212\u03d53X = \u03bda2F D(2\u03bd2 + \u03bd \u2212 3)b , \u03d52Y = \u03d53Y = (\u03bd2 + \u03bd \u2212 3)aF D(3 + 2\u03bd)(\u22121 + \u03bd) . The analytical solution is obtained as uz,max = \u2212 16Fa2 Eh3 for theEuler\u2013Bernoulli beam and as uz,max = \u2212 16Fa2 Eh3 \u2212 6F 5hG for the Timoshenko beam. It should be noted here that the Euler\u2013Bernoulli solution is obtained as a special case from u2Z for \u03bd \u2192 0. 6.2 Example: Four-element Example of a Plate Fixed at all Four Edges Given is a classical plate which is fixed at all four sides, see Fig. 6.9. The side lengths are equal to 4a. The plate is loaded by a single forces F acting in the middle of the plate. The material is described based on the engineering constants Young\u2019s modulus E and Poisson\u2019s ratio \u03bd. Use four classical plate elements (each 2a \u00d7 2a \u00d7 h) in the following to model the problem and to calculate the nodal unknowns in the middle of the plate. 308 6 Classical Plate Elements 6.2 Solution The elemental stiffnessmatrix of a classical plate elementwith dimensions 2a\u00d72a\u00d7h can be taken from Example6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure6.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure6.5-1.png",
        "caption": "Fig. 6.5 Rectangular four-node plate element: a Cartesian and b parametric space",
        "texts": [
            "62) 292 6 Classical Plate Elements which is a (3\u00d7n)-matrix. The last relation can be alsowritten directly for theB-matrix as: B = (L2NT)T = NLT 2 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 N1 N2 ... Nn \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 [ \u22022 \u2202x2 \u22022 \u2202y2 2 \u22022 \u2202x\u2202y ] (6.63) = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 \u22022N1 \u2202x2 \u22022N1 \u2202y2 2 \u22022N1 \u2202x\u2202y \u22022N2 \u2202x2 \u22022N2 \u2202y2 2 \u22022N2 \u2202x\u2202y ... ... ... \u22022Nn \u2202x2 \u22022Nn \u2202y2 2 \u22022Nn \u2202x\u2202y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 . (6.64) 6.3.2 Rectangular Four-Node Plate Element A simple representative of a two-dimensional plate5 is a rectangular four-node element (also called \u2018quad 4\u2019) as shown in Fig. 6.5, see [34, 35, 14]. The node numbering must follow the right-hand convention as indicated in the figure. Interpolation Functions and Derivatives Let us assume in the following a fourth-order polynomial for the displacement field ue z(\u03be, \u03b7) in the parametric \u03be-\u03b7 space: ue z(\u03be, \u03b7) = a1 + a2\u03be + a3\u03b7 + a4\u03be 2 + a5\u03be\u03b7 + a6\u03b7 2 + a7\u03be 3 + a8\u03be 2\u03b7 + a9\u03be\u03b7 2 + a10\u03b7 3 + a11\u03be 3\u03b7 + a12\u03be\u03b7 3, (6.65) or in vector notation 5An excellent review of classical plate elements is given in [29]. 6.3 Finite Element Solution 293 ue z(\u03be, \u03b7) = \u03c7Ta = [ 1 \u03be \u03b7 \u03be2 \u03be\u03b7 \u03b72 \u03be3 \u03be2\u03b7 \u03be\u03b72 \u03b73 \u03be3\u03b7 \u03be\u03b73 ] \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 a1 a2 ... a11 a12 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 . (6.66) Differentiation with respect to the y- and x-coordinate gives the rotational fields as (see Fig. 6.5): \u03d5e x(\u03be, \u03b7) = \u2202ue z \u2202y = \u2202ue z(\u03be, \u03b7) \u2202\u03b7 \u2202\u03b7 \u2202y = 1 b \u2202ue z(\u03be, \u03b7) \u2202\u03b7 (6.67) = 1 b [ 0 0 1 0 \u03be 2\u03b7 0 \u03be2 2\u03be\u03b7 3\u03b72 \u03be3 3\u03be\u03b72 ] \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 a1 a2 ... a11 a12 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 , (6.68) or \u03d5e y(\u03be, \u03b7) = \u2212\u2202ue z \u2202x = \u2212\u2202ue z(\u03be, \u03b7) \u2202\u03be \u2202\u03be \u2202x = \u22121 a \u2202ue z(\u03be, \u03b7) \u2202\u03be (6.69) = 1 a [ 0 \u2212 1 0 \u2212 2\u03be \u2212 \u03b7 0 \u2212 3\u03be2 \u2212 2\u03be\u03b7 \u2212 \u03b72 0 \u2212 3\u03be2\u03b7 \u2212 \u03b73 ] \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 a1 a2 ... a11 a12 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 . (6.70) Equations (6.66), (6.68) and (6.70) can bewritten inmatrix form for all four nodes as: \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 ue1z \u03d5e 1x \u03d5e 1y ue2z \u03d5e 2x \u03d5e 2y ue3z \u03d5e 3x \u03d5e 3y ue4z \u03d5e 4x \u03d5e 4y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \ufe38 \ufe37\ufe37 \ufe38 up = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 1 \u03be1 \u03b71 \u03be21 \u03be1\u03b71 \u03b721 1 b (0 0 1 0 \u03be1 2\u03b71 1 a (0 \u22121 0 \u22122\u03be1 \u2212\u03b71 0 \u03be31 \u03be21\u03b71 \u03be1\u03b7 2 1 \u03b731 \u03be31\u03b71 \u03be1\u03b7 3 1 0 \u03be21 2\u03be1\u03b71 3\u03b721 \u03be31 3\u03be1\u03b7 2 1 ) \u22123\u03be21 \u22122\u03be1\u03b71 \u2212\u03b721 0 \u22123\u03be21\u03b71 \u2212\u03b731 ) 1 \u03be2 \u03b72 \u03be22 \u03be2\u03b72 \u03b722 1 b (0 0 1 0 \u03be2 2\u03b72 1 a (0 \u22121 0 \u22122\u03be2 \u2212\u03b72 0 \u03be32 \u03be22\u03b72 \u03be2\u03b7 2 2 \u03b732 \u03be32\u03b72 \u03be2\u03b7 3 2 0 \u03be22 2\u03be2\u03b72 3\u03b722 \u03be32 3\u03be2\u03b7 2 2 ) \u22123\u03be22 \u22122\u03be2\u03b72 \u2212\u03b722 0 \u22123\u03be22\u03b72 \u2212\u03b732 ) 1 \u03be3 \u03b73 \u03be23 \u03be3\u03b73 \u03b723 1 b (0 0 1 0 \u03be3 2\u03b73 1 a (0 \u22121 0 \u22122\u03be3 \u2212\u03b73 0 \u03be33 \u03be23\u03b73 \u03be3\u03b7 2 3 \u03b733 \u03be33\u03b73 \u03be3\u03b7 3 3 0 \u03be23 2\u03be3\u03b73 3\u03b723 \u03be33 3\u03be3\u03b7 2 3 ) \u22123\u03be23 \u22122\u03be3\u03b73 \u2212\u03b723 0 \u22123\u03be23\u03b73 \u2212\u03b733 ) 1 \u03be4 \u03b74 \u03be24 \u03be4\u03b74 \u03b724 1 b (0 0 1 0 \u03be4 2\u03b74 1 a (0 \u22121 0 \u22122\u03be4 \u2212\u03b74 0 \u03be34 \u03be24\u03b74 \u03be4\u03b7 2 4 \u03b734 \u03be34\u03b74 \u03be4\u03b7 3 4 0 \u03be24 2\u03be4\u03b74 3\u03b724 \u03be34 3\u03be4\u03b7 2 4 ) \u22123\u03be24 \u22122\u03be4\u03b74 \u2212\u03b724 0 \u22123\u03be24\u03b74 \u2212\u03b734 ) \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \ufe38 \ufe37\ufe37 \ufe38 \u03c7 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 a1 a2 a3 a4 a5 a6 a7 a8 a9 a10 a11 a12 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \ufe38 \ufe37\ufe37 \ufe38 a , (6.71) 294 6 Classical Plate Elements where \u03bei and \u03b7i are the nodal coordinates in the \u03be-\u03b7 space, see Fig. 6.5b. Solving for a under consideration of these coordinates gives: \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 a1 a2 a3 a4 a5 a6 a7 a8 a9 a10 a11 a12 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 1 4 b 8 \u2212 a 8 1 4 b 8 a 8 1 4 \u2212 b 8 a 8 1 4 \u2212 b 8 \u2212 a 8 \u2212 3 8 \u2212 b 8 a 8 3 8 b 8 a 8 3 8 \u2212 b 8 a 8 \u2212 3 8 b 8 a 8 \u2212 3 8 \u2212 b 8 a 8 \u2212 3 8 \u2212 b 8 \u2212 a 8 3 8 \u2212 b 8 a 8 3 8 \u2212 b 8 \u2212 a 8 0 0 a 8 0 0 \u2212 a 8 0 0 \u2212 a 8 0 0 a 8 1 2 b 8 \u2212 a 8 \u2212 1 2 \u2212 b 8 \u2212 a 8 1 2 \u2212 b 8 a 8 \u2212 1 2 b 8 a 8 0 \u2212 b 8 0 0 \u2212 b 8 0 0 b 8 0 0 b 8 0 1 8 0 \u2212 a 8 \u2212 1 8 0 \u2212 a 8 \u2212 1 8 0 \u2212 a 8 1 8 0 \u2212 a 8 0 0 \u2212 a 8 0 0 a 8 0 0 \u2212 a 8 0 0 a 8 0 b 8 0 0 \u2212 b 8 0 0 b 8 0 0 \u2212 b 8 0 1 8 b 8 0 1 8 b 8 0 \u2212 1 8 b 8 0 \u2212 1 8 b 8 0 \u2212 1 8 0 a 8 1 8 0 a 8 \u2212 1 8 0 \u2212 a 8 1 8 0 \u2212 a 8 \u2212 1 8 \u2212 b 8 0 1 8 b 8 0 \u2212 1 8 b 8 0 1 8 \u2212 b 8 0 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 ue 1z \u03d5e 1x \u03d5e 1y ue 2z \u03d5e 2x \u03d5e 2y ue 3z \u03d5e 3x \u03d5e 3y ue 4z \u03d5e 4x \u03d5e 4y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 , (6",
            "319) \u03d52X = \u2212\u03d53X = \u03bda3F 2D(4a2 + 2b2 \u2212 a2\u03bd2 \u2212 2b2\u03bd) , (E.320) \u03d52Y = \u03d53Y = 0. (E.321) The special case \u03bd \u2192 0 gives uZ = \u2212 2a3F Ebh3 and \u03d5X = \u03d5Y = 0 which is equal to the Euler\u2013Bernoulli solution. 6.9 Symmetry solution for a plate fixed at all four edges Reduced system of equations: K7-7 \u00d7 u3Z = \u2212 F 4 . (E.322) Solution: u3Z = 10a2 D(27 \u2212 2\u03bd) \u00d7 F 4 = \u2212 5a2F 2D(27 \u2212 2\u03bd) . (E.323) 6.10 Investigation of displacement and slope consistency along boundaries Consider the boundary (x, y = 0), i.e. between node 1 and 2 in Fig. 6.5a. Evaluation of Eqs. (6.65), (6.68) and (6.70) in Cartesian coordinates for y = 0 gives: 474 Appendix E: Answers to Supplementary Problems ue z(y = 0) = a1 + a2x + a4x2 + a7x7, (E.324) \u03d5e x (y = 0) = a3 + a5x + a8x2 + a11x3, (E.325) \u03d5e y(y = 0) = \u2212(a2 + 2a4x + 3a7x2). (E.326) Four DOF from node 1 and 2 can be used to determine a1, a2, a4 and a7 and thus ue z and \u03d5e y (which are continuous along element boundaries). However, the remaining two DOF do not allow to uniquely define the four constants a3, a5, a8 and a11 for \u03d5e x "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000851_amm.658.89-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000851_amm.658.89-Figure1-1.png",
        "caption": "Fig. 1. Dynamic model with six degrees of freedom (6 DOF) [3]",
        "texts": [
            " The final scopes of mathematical modeling of gears could be summarized as follows: - Analysis of contact and bending stress; - Reduction of superficial wear as for example pitting; - Study of transmission efficiency; - Study of noise radiation; - Natural frequencies of the system; - Studies regarding the vibratory motion of the system. The mathematical models for the dynamic simulation of gearboxes, proposed by several investigators [1,2], are showing considerable variations, not only in the level of the basic assumptions, but also in the effects. When the influence of the gear carrying shafts, of the bearings and the driving, respective driven machine cannot be neglected, a mathematical model with six degrees of freedom (6 DOF) has to be chosen. Such a model is shown in Fig. 1. The equation of motion for the above dynamic system can be written in following form: [ ] { } [ ] { } [ ] { } { }FqCqDqM =\u22c5+\u22c5+\u22c5 , (1) where: [M]- Matrix of masses, [D]- Matrix of damping\u2019s, [C]- Matrix of stiffness, q - vector of deplacements, and F -vector of force. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 149.171.67.164, University of New South Wales, Sydney, Australia-08/07/15,10:34:53) The terms of equation (1) can be written, according to [3], as follows: [ ]                                   \u22c5 \u22c5 \u22c5 \u22c5 = 2 1 2 1 2 1 2 2 1 1 2 1 00000 00000 00 4 000 000 4 00 0000 4 0 00000 4 1 m m r J r J r J r J m M b Ma b b b Mo red , (2) [ ]                                       + \u2212+\u2212\u2212 \u22c5 \u22c5 \u22c5 \u22c5 \u22c5 \u2212 \u22c5 \u2212        \u22c5 \u22c5 + \u22c5 \u2212 \u22c5 \u22c5 \u2212        + \u22c5 \u22c5 \u2212 \u22c5 \u22c5 = z zz b Ma b bz b bz b Ma b bz b Ma zz b bzz b Mo b Mo b Mo redz dd dddsymmetric r D r rd r rd r D r rd r D dd r rdd r D r D r D mc D 2 1 2 1 1 2 1 2 2 11 2 2 1 1 2 2 1 2 1 2 1 00 4 22 4 4 4 22 0 44 4 0000 44 1 , (3) [ ]                                         + \u2212+\u2212\u2212 \u22c5 \u22c5 \u22c5 \u22c5 \u22c5 \u2212 \u22c5 \u2212        \u22c5 \u22c5 + \u22c5 \u2212 \u22c5 \u22c5 \u2212         + \u22c5 \u22c5 \u2212 \u22c5 = z zz b Ma b bz b bz b Ma b bz b Ma zz b bzz b Mo b Mo b Mo z cc cccsymmetric r C r rc r rc r C r rc r C cc r rcc r C r C r C c C 2 1 2 1 1 2 1 2 2 11 2 2 1 1 2 2 1 2 1 2 1 00 4 22 4 4 4 22 0 44 4 0000 44 1 , (4) { }T MaMo yyq 2121 \u03d5\u03d5\u03d5\u03d5= . (5) For this purpose, the facilities of the SolidWorks software have been used. The 3D geometry and the finite element discretization of the gearbox, having the technical data as described in table 1, are shown in Fig. 2. The dynamic simulation of the gearbox has been performed using the Linear Dynamic Analysis tool of the software, respective the Modal Time History Analysis. For the dynamic simulation there has been used the 6 DOF mathematical model shown in Fig. 1, motion being described by Eq. 1. As it is well known, one of the main sources of gear vibration is the variation of the tooth force along the length of engagement. Assuming a tooth force variation according to [4] pp. 101, for the six loading steps (Mt2= 22,0; 26,0; 29,5; 32,5; 35,5; 38,5 [Nm]) has been calculated the resulting accelerations (above the bearing of the high speed shaft from the motor side), results being presented in table 2. Fig. 3 shows the resulting accelerations calculated for Mt2= 38,5 Nm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure7.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure7.7-1.png",
        "caption": "Fig. 7.7 Regime change at the expense of conductivity y0N ! y0N",
        "texts": [
            " b Conductance yiN presets current IG11 218 7 Recalculation of Load Currents of Active Multi-ports Internal conductivity (7.13) yiN \u00bc 2:5\u00fe 0:25 1:25 0:25\u00fe 1:25 \u00fe 0:25 0:833 0:25\u00fe 0:833 \u00bc 2:9 : We check transformation parameter (7.14) mN \u00bc yN \u00fe yiN yN \u00fe yiN \u00bc 0:625\u00fe 2:9 1:25\u00fe 2:9 \u00bc 0:8494 : 7.2.2 Change of Longitudinal Conductivity We consider again the circuit in Fig. 7.2. Let the conductivity y0N be changed, y0N ! y0N . Corresponding changes of an initial pointM ! M and short circuit point SC ! SC are shown in Fig. 7.7. Also, the coordinates of points G1; G2 by (6.24) and (6.26) are proportional to the value y0N . Therefore, the subsequent straight line G1 G2 is parallel to the initial line G1 G2; the values YG1 L1 ; Y G2 L2 do not change. Let us determine fixed points and lines as y0N change. Naturally, the point 0 does not depend on this element. If the current across conductivity y0N is equal to zero, a straight line S1 S2 is the fixed line and we have the following equation of this line: y1 \u00fe y1N 7.2 Recalculation of Currents for the Case of Changes of Circuit Parameters 219 The internal conductivity yi0N of the circuit relatively to terminals of conductivity y0N , by Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002084_detc2015-47180-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002084_detc2015-47180-Figure1-1.png",
        "caption": "Figure 1. EXPERIMENTAL VEHICLE ON THE BELGIAN PAVING",
        "texts": [
            " BS British Standards Institution DGPS Differential Global Positioning System DSD Displacement Spectral Density FFT Fast Fourier Transform ISO International Organization for Standardization OEM Original Equipment Manufacturer rms Root mean square 4S4 Four state semi-active suspension system 1 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 08/12/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use EXPERIMENTAL VEHICLE The experimental vehicle is a 1997 Land Rover Defender 110 Station Wagon (see Fig.1). The OEM suspension has been removed and replaced with a hydro-pneumatic suspension. The hydro-pneumatic suspension consists of two gas accumulators (0.1l and 0.4l) that are separated from the hydraulic fluid with two floating pistons. The suspension can switch between using either both the gas accumulators, or only a single one, resulting in either soft or stiff spring characteristics. The hydraulic fluid passes through a valve manifold that contains two damper packs fitted with bypass valves"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002748_ijmic.2016.075271-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002748_ijmic.2016.075271-Figure1-1.png",
        "caption": "Figure 1 Inertial reference frame {I} and body coordinate frame {B}",
        "texts": [
            "1 Notations I is the identity matrix of appropriate dimension. AT denotes transpose of matrix A, The norm of vector x is defined as || || Tx x x= and || || ( )TA \u03bb A A= a is norm of matrix A where \u03bb is the maximum eigenvalue of matrix ATA. In the paper, \u03bb and \u03bb are used to indicate largest and smallest eigenvalues respectively. 2 AUV model and properties The three-dimensional equations of motion for AUV were described by Fosen (1994) using inertial reference frame {I} and body fixed frame {B} as shown schematically in Figure 1. As the rotation of the Earth has little effect on low speed underwater vehicles, Earth fixed reference frame can be considered as an inertial reference frame. 2.1 Kinematics The linear and angular velocities in body fixed frame {B} are related to the velocities in inertial reference frame {I} using Jacobian matrix R(\u03b72) as ( ) ( ) ( )1 11 2 2 2 22 2 0 0 \u03b7 vJ \u03b7 \u03b7 R \u03b7 v \u03b7 vJ \u03b7 \u23a1 \u23a4\u23a1 \u23a4 \u23a1 \u23a4 = \u21d2 =\u23a2 \u23a5\u23a2 \u23a5 \u23a2 \u23a5 \u23a3 \u23a6 \u23a3 \u23a6\u23a3 \u23a6 (1) where \u03b71 = [x y z]T and \u03b72 = [\u03c6 \u03b8 \u03c8]T represent position and orientation of the AUV respectively, expressed in inertial reference frame"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001145_ijde.2014.062378-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001145_ijde.2014.062378-Figure1-1.png",
        "caption": "Figure 1 Input dimensions of CP-CVT",
        "texts": [
            " Everything within the drive itself can be mechanically automated and hence requires no external hydraulic or servo control mechanism. The design itself, which has been described in several different papers (Cretu and Glovnea, 2005, 2006; Bell et al., 2011) consists of a conical input disc, a conical output disc, a toroidal input disc and a convenient number of intermediary spherical elements. The key components can hence be entirely described through five input dimensions, two angular (\u03b2 and \u03b3) and three linear (R, R1, r0), as shown in Figure 1. The output shaft is connected to the conical output disc through a ball screw coupling, hence any resistive torque applied to the output shaft will cause an axial load (FC) to be applied to the ball element at C, displacing it in radial direction. The contact force between the balls and the toroidal disc will be balanced by a force FA at point A, produced by a linear spring. By carefully designing the elasticity of the spring and the geometric characteristics of the ball-screw coupling the position of the ball element is adjusted automatically",
            " Hence, for this problem, seeding was favoured over mutation: instead of randomly changing a single bit value, occasionally a completely new chromosome was added to a generation replacing an existing weaker one. This \u2018seeding\u2019 is applied stochastically to prevent any strong solutions being prematurely removed from a generation and to ensure that the new seed was immediately feasible. In addition to reducing the probability of extinct bits, this also had the advantage of ensuring that generations maintain a healthy population of feasible solutions. A basic set of design constraints can be inferred from Figure 1. Although more advance constraints were previously used (Bell, 2011), this basic set of rules is computationally efficient and was shown to be sufficient for the vast majority of solutions. These constraints are required to ensure that the evaluated solutions are both physically possible to construct and mathematically calculable: 1 \u03b2 must be smaller than \u03b3, else the ball element cannot be in equilibrium 2 R1 must be larger than R, or the surface of the ball elements will be larger than the curvature of the toroidal disc 3 r0 must be larger than R1, or the curvature of the toroidal disc will be interfere with the input shaft"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000218_978-3-319-52219-7-Figure5.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000218_978-3-319-52219-7-Figure5.5-1.png",
        "caption": "Fig. 5.5 (a), (b) Design and structure detail of the electronically controlled THz metamaterial absorber based spatial light modulator (MMA-SLM) [25]. (a) Image of MMA-SLM as assembled in chip carrier package. (b) Cross-sectional schematic view of a single pixel. (c) Schematic of a THz compressive imaging setup using the SLM in (a), (b) [16]",
        "texts": [
            " The principle of operation lies in the fact that CW pumped -Si behaves as THz modulator due to free carrier generation and linear recombination in the semiconductor changing the complex dielectric constant that can be described by the Drude model [27] .!/ D 1 !2 p !.! C i / (5.3) 88 5 A Terahertz Spatial Light Modulator for Imaging Application A fundamental limitation of this technique is the carrier lifetime of -Si of D 25 s limits the switching speed to about 10 kHz. Moreover, the LED source, DMD, and the optics to create the spatial CW pumping make the system complex and expensive. Figure 5.5a, b shows the SLM based on multi-resonant, electronically controlled metamaterial absorber [25]. The overall SLM system architecture is shown schematically in Fig. 5.5b and consists of metamaterial absorber pixels flip chip bonded to a Silicon chip carrier with routing to bond pads which are wirebonded to a leadless chip carrier (LCC). The metamaterial absorber consists of two metallic layers with a dielectric spacer lying in-between. The top metal layer is patterned in order to respond resonantly to the electric component of an incident 5.2 A Review of THz Spatial Light Modulators 89 electromagnetic wave. A bottom ground plane layer is spaced relatively close to the top layer, thus allowing the external magnetic field to couple, as shown in Fig. 5.5b. The SLM was demonstrated in a single-pixel compressive imaging experiment using a reflection geometry [16], a schematic of the setup shown in Fig. 5.5c. The same SLM has also been used in recent works involving single-pixel quadrature imaging [28] and frequency-division-multiplexed single-pixel imaging [29]. This electrically controlled SLM is a great improvement over the optically pumped SLM allowing for simpler and cheaper THz imagers. However, having to use a certain height of dielectric to design for a desired resonant frequency for the absorber makes the design very rigid and precludes the use of commercial foundry process where the vertical dimensions are fixed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000618_dac.2954-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000618_dac.2954-Figure1-1.png",
        "caption": "Figure 1. A priority event area in the sense field.",
        "texts": [
            " The priority function is also shown as \u03c6(q) = exp( \u03b5 \u00b7 [(x xq) 2 + (y yq) 2]). To obtain better coverage performance, a distributed deployment strategy is presented in this paper. That is, in regions where the value of \u03c6(q) is high, the density of sensors deployed is correspondingly greater. Using this deployment, the data sensed by the nodes is more reliable. Sometimes, the measure used represents information of the probability that some event takes place over A. A priority event of area in sense field is shown in Figure 1. A sensing error function is proposed to estimate the effectiveness of sensor nodes deployment. The sensing error of a sensor is directly related to the distance between the point and the sensor. Due to noise and loss of resolution, the sensing performance at a point q taken from the i-th sensor (position pi) degrades with the distance \u2016q pi\u2016 between q and pi. We describe this degradation using a non-decreasing differentiable function f. Accordingly, f(\u2016q pi\u2016) provides a quantitative assessment for sensing performance at different points"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.3-1.png",
        "caption": "Fig. 4.3 a Four-bar mechanism; b triple pendulum",
        "texts": [
            " The number of degrees-of-freedom of a multibody system can be evaluated as the difference between the system coordinates and the number of independent constraints or relationships among these coordinates. The mathematical expression that summarizes this concept for planar motion is known as the Gr\u00fcebler-Kutzback criterion and is written as (Shigley and Uicker 1995) DOF \u00bc 3nb m \u00f04:1\u00de where nb represents the number of rigid bodies that compose the multibody system and m is the number of independent constraints. For example, the planar four-bar mechanism shown in Fig. 4.3a has four bodies (including ground), four revolute joints with two constraints for each revolute joint, and three ground body constraints, yielding one degree-of-freedom. Figure 4.3b shows a triple pendulum that comprises four bodies, six revolute joint constraints for the three revolute joints, and three ground body constraints, which results in three degrees-of-freedom. It is not unanimous and it is not a simple task either to define a criterion to classify the different types of coordinates that can be used to describe the configuration of multibody systems. A general and broad embracing rule to group the generalized coordinates is to divide them into \u201cindependent\u201d and \u201cdependent\u201d coordinates (Wehage and Haug 1982)",
            " Therefore, Eulerian coordinates require that a large number of coordinates be defined to specify the position of each body of a multibody system. In the present work, the vector of coordinates q is defined as a column vector that contains all the variables used in the description of the configuration of multibody systems. In the following paragraphs, the different types of coordinates are briefly presented by using simple demonstrative examples. Let consider the triple pendulum illustrated in Fig. 4.3b, from which it can be observed that the configuration of the system can uniquely and completely be defined if the angular variables \u03d51, \u03d52 and \u03d53 are known for each instant of time during the analysis. This is obvious as the triple pendulum has three degrees-of-freedom in the measure that it has three independent motions, each one associated with each arm rotation. In this case, the set of generalized (or independent) coordinates can be expressed as q \u00bc /1 /2 /3f gT \u00f04:2\u00de where the superscript T represents the transpose mathematical operation. In a similar way, in the four-bar mechanism presented in Fig. 4.3a there is a set of three angular variables \u03d51, \u03d52 and \u03d53 that define the configuration of the system. However, these three variables are not independent because the system has only one degree-of-freedom. The angular variables can be related to each other by writing two algebraic equations of the closed kinematic chain associated with the four-bar mechanism, yielding a cos/1 \u00fe b cos/2 c cos/3 d \u00bc 0 \u00f04:3\u00de a sin/1 \u00fe b sin/2 c sin/3 \u00bc 0 \u00f04:4\u00de in which a, b, c and d are the lengths of the links. Thus, Eqs",
            "8 results that a cos/1 \u00fe b cos/2 \u00fe c cos/3 d \u00bc 0 \u00f04:16\u00de a sin/1 \u00fe b sin/2 \u00fe c sin/3 \u00bc 0 \u00f04:17\u00de Thus, for a given configuration, that is, known for instance \u03d51, the set algebraic equations (4.16) and (4.17) must be solved simultaneously for \u03d52 and \u03d53. This procedure is generally preformed numerically. In summary, the relative coordinates are used to formulate a minimum number of equations of motion of multibody systems. When the system is an open kinematic chain, the number of relative coordinates is equal to the number of degrees-of-freedom, as it is the example of the triple pendulum illustrated in Fig. 4.3b. In these circumstances, the relative coordinates are in fact the independent variables used to define the configuration of the system. For closed kinematic chains, a preprocessing analysis of the system is required to deal with the assembling constraints, and then the system topology has to be analyzed to describe the constants properly. Therefore, relative coordinates are not convenient when the system topology can be altered during the global motion produced. Furthermore, in sharp contrast to the absolute coordinates, the incorporation of general force functions, constraint equations and prescribed trajectories in the system\u2019s formulation are not trivial tasks when relative coordinates are used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002373_21681015.2016.1174162-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002373_21681015.2016.1174162-Figure1-1.png",
        "caption": "Figure 1.\u00a0sketch of the configuration and generation of complex cycloid tooth profile (tP).",
        "texts": [
            " As the conjugate features of flexspline have been researched before [19], in this paper, the TP of circular spline is obtained by envelope method, and to make this TP expressed well, it was fitted by Complex Method. The objective function was set as the summation of the distances between the points and the fitted curves. The optimal solution according to the Complex Method was used as parameters of a circular spline TP to build an assemble model with the flexspline, on which the backlashes between the engaged TPs were measured in the visualization environment. Complex TP is composed of an epicycloid and a hypocycloid. Its configuration and generation are shown in Figure 1. O1 is the origin of coordinate system {O1, X1, Y1} that is located on the neutral line of the flexspline. O is the origin of coordinate system {O, X2, Y2} that is located on the rotation axis of the flexspline. Axis Y1 coincides with the symmetry line of the tooth, and R is the radius of the reference circle. The epicycloid profile of tooth face, curve AB, is created by the trajectory of a point on the outside circle with radius rw rolling outside the reference circle. The hypocycloid profile of the tooth flank, curve BC, is created by the trajectory of a point on the inside circle with radius as rn, rolling inside the reference circle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001271_amm.611.279-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001271_amm.611.279-Figure7-1.png",
        "caption": "Fig. 7 Sample solutions to stress in gear by FEM",
        "texts": [],
        "surrounding_texts": [
            "This problem is solved for gear with variable transmission in the range u = 0,5 to 2,0 ,with the number of teeth z1 = z2 = 24 and gearing module mn = 3,75 mm, the axial distance a = 90 mm and for a one sense of rotation. To create this gear is analyzed in detail in the literature [5] and [6]. The gears for a given variable transmission have been proposed as elliptical - eccentrically placed (Fig. 2), so that conditions were right shot. Real of load gear teeth with variable gear ratio is not constant. By way of illustration is given unit input torque (driven) spur gear Mk1 = 100Nm. In Figure 3 the course of torque Mk1 on the input gear and torque Mk2 on the output (driven) gear (Mk2i = Mk1.ui) is show. In Figure 4 are value of changing tangential tooth load the driver and driven gear F01 = F02 if F01=Mk1/r1i, radial force Fr1 = Fr2 if Fr1 = F01.tg\u03b1 . The resultant force acting on the side of the tooth FN1 = FN2 if FN1 = F01 / cos\u03b1, where \u03b1 is an angle of action to 20\u00b0. A B F bn 30\u00b0 F bn 30\u00b0 snF X Y Fig. 5 The tooth load Kinematic conditions were processed for a gear 1 (the center of rotation at point O1) and the gear 2 (with the center of rotation at point O2). Stress of teeth in dangerous section solution by FEM Create a geometric model of the gear is the first step to deal with tooth stress by FEM. Universal instructions to create geometry computer model does not exist [7]. The first part was to develop a functional model gear designated for the production of gears gearing for NC machine to electrospark cutting. It is this suitably modified dxf format describing the shape of gears was used to create a geometric model. To determine the computer model for studying deformation of the teeth using FEM was necessary to determine the material constants, define the type of finite element, and selecting appropriate boundary conditions (geometry and power). To determine the stress at the foot of the tooth it is necessary to know the distribution of load on individual pairs of teeth in the meshing. To start with let us consider the simplest, the ideal distribution of the load when the load-pair meshing are divided in half for each pair of meshing. The problem is solved with the gear continuously variable transmission numbers. The stress in a dangerous section of the tooth is solved using the finite element method for driving gear, the gear teeth to reach the number 0.5, 1 and 2. In Figure 5 the tooth load is shown. The resulting stress in dangerous cross-section is considered at points X and Y after the width of the teeth. It is not necessary to solve this problem in a model full the wheel of gear. In Figure 6 the segments of gear are shows. In Table 1 are results of stress in the dangerous section of tooth solution by finite element method for segments (Fig. 6) driver elliptical gear set with continuously variable gear ratio. Width of teeth is 10 mm, the driving torque is Mk1=100 Nm. gear ratio u = 1 (Fig. 6) under load according to Figure 5. In Figure 8 the representation of the medium stress in a dangerous section of tooth gear segments for the gear ratio u = 0,5, u = 1 and = 2 (Fig. 6) is shown. The stress at the load of the tooth point A on the side load force (Fig. 5) has be specification A-X. The results show that the stress in a dangerous section of teeth on the load side and on the opposite side (at point Y and point X Fig. 5) is different. The stress in the foot the tooth drive gear increases with gear ratio. Fig.8 The medium stress in the tooth foot"
        ]
    },
    {
        "image_filename": "designv11_64_0002610_1.4034343-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002610_1.4034343-Figure1-1.png",
        "caption": "Fig. 1 Schematic example of bump-type first-generation foil gas bearings with axially and circumferentially uniform elastic support elements [1]",
        "texts": [
            " The data exchange is done for every iteration step, so this FSI model is fully coupled, typically referred to as two-way coupled. The model is validated with the experimental test data by Ruscitto et al. [18]. Foil gas bearing load capacity and attitude angle are the main performance parameters of interested in the investigation. The FSI model\u2019s predictions show good agreement with test data. The foil gas bearing\u2019s (FGB\u2019s) geometry studied was based on the bearing in Ref. [18], and geometric details are shown in Table 1. This bearing is a bump-type first-generation foil gas bearing as shown in Fig. 1. The top and the single bump foils are spot welded at one end to the bearing sleeve, while the other end is free to deform in the circumferential direction. The journal rotational direction is from free end to fixed end as shown in Fig. 1. The tests were performed under ambient condition in Ref. [18] and are used as validation data for the fully coupled FSI model in this paper. The authors report a nominal radial clearance of 31.8 lm based on an ad hoc procedure displacing the journal with Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received August 11, 2015; final manuscript received August 28, 2015; published online September 13, 2016",
            " The authors have modeled bearings with higher eccentricity values, which provide higher static loads, and extend the predictions\u2019 lines to the right side but computational effort does increase; these model results are not included in this paper. Figure 7 presents journal attitude angle versus static load for the 30 krpm, 45 krpm, and 55 krpm conditions for a bearing L/ D\u00bc 1. The attitude angle is evaluated as the included angle for FX and FY, which refer to the forces in the X and Y direction in the bearing; X and Y are the horizontal and vertical directions, respectively, in Fig. 1, for example. The graph compares test data [18] and FSI model predictions. The FSI model predictions show good agreements with test data when the static load is larger than 40 N for bearing speeds of 30 krpm and 45 krpm. The predicted results slightly underestimate the test data for condition 55 krpm. All predictions indicate a static load equal to 0 at about 90 deg. Figure 8 presents minimum film thickness versus achieved static load at the midplane for rotational speeds at 30 krpm, 45 krpm, and 55 krpm for the bearing with L/D\u00bc 1/2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002200_naps.2014.6965444-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002200_naps.2014.6965444-Figure3-1.png",
        "caption": "Fig. 3. Selection of a voltage vector in sector 1; (a) voltage vector v1 to intensify and (b) voltage vector v2 to weaken the stator flux vector.",
        "texts": [
            " Both variables are functions of the angle between the stator flux vector and D-axis (i.e., \u03b8). The amplitude of the four voltage vectors of a two-leg VSC are the same. Thus, the time interval of applying any of those voltage vectors is proportional to the path length that the stator flux traverses as long as the stator flux modulus is in the allowed range. In the first case, it is assumed that the stator flux modulus has reached to the minimum allowed value ( |\u03bb\u0304s|\u2217 \u2212\u2206\u03bb ) in the sector 1. From Fig. 3(a), v1 is selected to increase the stator flux modulus. In Fig. 3(a), |\u03bb\u0304s|\u2217 is the desired stator flux modulus, \u2206\u03bb is the width of the hysteresis controller, and x is the length of the flux-developing vector. In OAB triangle, (9) is established:( |\u03bb\u0304s|\u2217 + \u2206\u03bb )2 = x2 + ( |\u03bb\u0304s|\u2217 \u2212\u2206\u03bb )2 \u2212 2x ( |\u03bb\u0304s|\u2217 \u2212\u2206\u03bb ) cos ( 3\u03c0 4 + \u03b8 ) . (9) Assuming k = \u2206\u03bb |\u03bb\u0304s|\u2217 and x\u2032 = x |\u03bb\u0304s|\u2217 , x\u2032 =\u221a 4k + (1 + k)2sin2 (\u03c0 4 + \u03b8 ) \u2212 (1 + k)sin (\u03c0 4 + \u03b8 ) . (10) Since k is a small value, Maclaurin series gives the normalized length of the flux-developing vector as x\u2032 = 2k sin ( \u03c0 4 + \u03b8 ) . (11) In the second case, it is assumed that the stator flux modulus has reached to the maximum allowed value ( |\u03bb\u0304s|\u2217 + \u2206\u03bb ) in the sector 1. From Fig. 3(b), v2 is selected to decrease the stator flux modulus. In Fig. 3(b), y is the length of the fluxweakening vector. Similar calculations in OA\u2032B\u2032 triangle, gives the normalized length of the flux-weakening vector as y\u2032 = 2k cos ( \u03c0 4 + \u03b8 ) . (12) Therefore, the probabilities of using the flux-developing and flux-weakening vectors can be derived as Pv1(\u03b8) = x\u2032 x\u2032 + y\u2032 = cos ( \u03c0 4 + \u03b8 ) sin ( \u03c0 4 + \u03b8 ) + cos ( \u03c0 4 + \u03b8 ) (13) Pv2(\u03b8) = y\u2032 x\u2032 + y\u2032 = sin ( \u03c0 4 + \u03b8 ) sin ( \u03c0 4 + \u03b8 ) + cos ( \u03c0 4 + \u03b8 ) . (14) From (8), (13), (14), and Table II, the linear speed of stator flux vector can be derived as Vp = Pv1(\u03b8)v1p(\u03b8) + Pv2(\u03b8)v2p(\u03b8) = Vdc 2cos(\u03b8) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003196_cistem.2014.7076970-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003196_cistem.2014.7076970-Figure6-1.png",
        "caption": "Fig. 6. Points de calcul de l'induction",
        "texts": [
            " Les bobines supraconductrices ont \u00e9t\u00e9 dimensionn\u00e9es en tenant compte de la loi de Jc (B) du fil supraconducteur. III. \u00c9SULTATS Nous allons v\u00e9rifier la n\u00e9cessit\u00e9 d'ajouter un mat\u00e9riau ferromagn\u00e9tique hyper satur\u00e9 entre les deux bobines supraconductrices autour de l'\u00e9cran, en comparant les courbes de l'induction radiale de la structure avec et sans fer. Les valeurs de la composante radiale de la densit\u00e9 de flux Br (Z) sont prises \u00e0 2 cm (pour tenir compte de la partie cryog\u00e9nique) de l'inducteur le long de la distance C entre les deux bobines. Les points de calculs sont pr\u00e9sent\u00e9s sur la Fig. 6. La composante radiale Br (\u03b8) est calcul\u00e9e \u00e9galement \u00e0 2 cm de l'inducteur sur le diam\u00e8tre d'al\u00e9sage de l'induit de distance pour plusieurs valeurs de Z entre les deux bobines. Ces calculs sont pr\u00e9sent\u00e9s dans les Fig. 8 (a, b et c). La Fig. 7 repr\u00e9sente l'\u00e9volution de Br le long de la direction Z pour la structure d'inducteur sans fer (ligne continue) et avec du fer (ligne avec des cercles). Nous atteignons avec du fer une valeur d'induction dans l'entrefer d'environ 1,9 T au maximum alors que nous n'atteignons que 1,7 T avec une structure sans fer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002813_b978-0-08-100072-4.00003-4-Figure3.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002813_b978-0-08-100072-4.00003-4-Figure3.5-1.png",
        "caption": "Figure 3.5 Cross section of boronate-affinity sensor for intravascular glucose monitoring showing the bornonate gel-covering dialysis membrane and outer haemocompatible, heparin-coated microporous membrane. With permission from Crane BC, Barwell NP, Gopal P, Gopichand M, Higgs T, James TD, et al. The development of a continuous intravascular glucose monitoring sensor. Journal of Diabetes Science and Technology 2015;9(4):751\u201361.",
        "texts": [
            " Alginate-entrapment gel was used, and though combined with biocompatible poly (vinyl alcohol) and a hydrophilic biocompatible coating, a tissue response still occurred that generated an avascular fibrous capsule. An intravascular fibreoptic probe has been developed using gel-loaded boronate. This is intended for short-term use in critically ill patients (\u223c40 h) [98]. The key materials goal is haemocompatibility, thus a dialysis membrane encloses the gel to serve as a barrier against proteins, close off this access route to glycated and non-glycated protein interference while a microporous membrane prevents ingress of blood cells (Fig. 3.5). The latter is also platinum loaded to destroy peroxides that are able to oxidise the boronate binder and combines an outer heparin layer to reduce initial blood cell adhesion. Sampling from the body is an alternative to sensor implantation, avoiding attendant tissue-interfacing concerns. The GlucoWatch system (Cygnus, California) attempted this [99]. Reverse iontophoresis was used to extract tissue fluid through skin, with sample taken up into an absorbent pad loaded with glucose oxidase, and formed from crosslinked PEG-polyacrylic acid"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.10-1.png",
        "caption": "Fig. 4.10 Four-bar mechanism described by natural coordinates natural coordinates",
        "texts": [
            " In the two-dimensional space, the natural coordinates can be seen as an extension of the absolute coordinates when the reference points are moved to relevant points of the multibody system. Figure 4.9 illustrates a schematic representation of this basic idea of the transition from absolute coordinates to natural coordinates. As it was mentioned, the main feature of the natural coordinates is that no angular variables are involved, but only Cartesian coordinates are used instead. Thus, the configuration of each body is described by at least two basic points properly located. For instance, in the case of the four-bar mechanism depicted in Fig. 4.10, the system can be described by four Cartesian coordinates x2, y2, x3 and y3, that is Since the four-bar mechanism has only one degree-of-freedom, then three constraint equations are required to describe the configuration of the system. The constraint equations must ensure that points 2 and 3 move according to the restrictions imposed on them by the three moving rigid bodies. Therefore, based on the concept of rigid links (defined by constant distances between points), the three following conditions can be written \u00f0x2 x1\u00de2 \u00fe\u00f0y2 y1\u00de2 a2 \u00bc 0 \u00f04:19\u00de \u00f0x3 x2\u00de2 \u00fe\u00f0y3 y2\u00de2 b2 \u00bc 0 \u00f04:20\u00de \u00f0x3 x4\u00de2 \u00fe\u00f0y3 y4\u00de2 c2 \u00bc 0 \u00f04:21\u00de As observed from Eqs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003602_1.5119067-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003602_1.5119067-Figure2-1.png",
        "caption": "Figure 2 Schematic of inside processing head for ultra-high-speed LMD with integrated continuous coaxial powder nozzle",
        "texts": [
            " The influence of the variation of the coating rate, powder mass flow and axial feed on the layer thickness is investigated. Coating rates are varied between 62.5 cm2/min and 200 cm2/min while powder mass flows are examined in the range of 15 g/min to 25 g/min. Inside processing head with integrated continuous coaxial powder nozzle Based on the basic principles of continuous coaxial powder feeding concepts for the ultra-high-speed LMD on outside diameters, a continuous coaxial powder feeding nozzle is developed and integrated into an inside LMD processing head from IXUN Lasertechnik [2], see Figure 2. The subassemblies, including beam guidance, beam forming and media supply (powder feed, shielding gas and purging gas for the protection of the optical components, water cooling circuit) are integrated into a compact cylindrical housing. The nozzle including the adaption to the tubus is designed modularly. With the processing head, work pieces with a length of up to approx. 1 m. and a minimal inside diameter of 100 mm can be processed. As beam source a 4kW Laserline LDF-4000-8 diode laser source with beam converter is linked to the inside processing head via a 200 \u00b5m fiber"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001292_0954406215582014-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001292_0954406215582014-Figure2-1.png",
        "caption": "Figure 2. Symmetric indentation of a flat surface by two rigidly interconnected wedge-shaped punches.",
        "texts": [
            " As an example, this analytical approach can be applied to the symmetric indentation of a flat surface by two wedge-shaped punches that their elastic properties are similar to the flat surface; provided the two punches are interconnected rigidly together and considered as one solid body indenting the lower half plane. These symmetric double contacts under the constant normal and oscillatory tangential loading are occasionally observed in fretting fatigue specimens tested to failure in laboratories. Example: Two rigidly interconnected wedge-shaped punches Figure 2 illustrates the symmetric indentation of a flat surface by two rigidly interconnected wedge-shaped punches with the small inner angles \u20191 and the small outer angles \u20192. 2 c is the horizontal distance between the two singular initial contact points. The loading Scenario is according to problem definition. By the half plane assumption, the mentioned approach is applied to analyze this problem. Utilizing equation (21), the completely nonsingular form of the pressure function during the first stage can be written as Ap\u00f0x\u00de \u00bc 2 xj j ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x2 a2\u00de\u00f0b2 x2\u00de q tan\u00f0 1\u00de K\u00f0x, s, a, b\u00de\u00f0 s\u00bc c s\u00bca tan\u00f0 2\u00de K\u00f0x, s, a, b\u00de\u00f0 s\u00bcb s\u00bc c \u00f029\u00de where K\u00f0x, s, a, b\u00de \u00bc R ds \u00f0x2 s2\u00de ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0b2 s2\u00de\u00f0s2 a2\u00de p which can be indicated as elliptic integrals; hence the pressure function of this problem cannot be expressed in terms of elementary functions of x"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure12-1.png",
        "caption": "Fig. 12. Details of the conductor car with the human body model in the vicinity of the insulators",
        "texts": [
            " ELECTRICAL SIMULATIONS Besides the mechanical analysis of the structure, evaluation of electrical aspects is also particularly important. For the inspection of electrical stresses, potential and electric field distribution has been examined in case of a critical geometry, where the clearances are close to the minimal distance required for safe live-line work in Hungary [2]. Fig. 11 shows a grounded tower structure and the phase conductors in the middle phase held by double composite insulators: it is a typical arrangement in the Hungarian 400 kV high voltage grid. Fig. 12 shows a closer picture of the simulated arrangement focusing on the model of the human body and the new conductor car. Human body model was created based on the guidance of IEC 62233 [3] with some modification required because of the sitting position of the worker during the simulations. In this case finite element solution of COMSOL MultiPhysics was used to determine both the electric potential and the electric field distribution in the vicinity of the arrangement. The first part of the simulation was the determination of the electric potential of the different elements of the arrangement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure2-1.png",
        "caption": "Fig. 2 Design domains of the caliper (green) and upright (red).",
        "texts": [
            " (2) instead provides the relation between the structural stiffness and the design variables given by the SIMP interpolation scheme. Starting from an initial solution (uniformly distributed material over the design domain), material distribution is iteratively updated until convergence to a minimum is reached. A finite element model of the brake caliper and the upright has been developed for topology optimization. The design and frozen domains referring to the caliper and the upright have been defined. The design domains of the brake caliper and the upright are depicted in Fig. 2. The frozen domain is shown in Fig. 3. The frozen domain is made by the hub bearing houses, the connections of the upright to the suspension wishbones and steering rod, by the connections between the caliper and the upright and by the six cylinders and pad supports in the brake caliper. This domain is fixed and is not involved in the optimization process. Design and frozen domains have been discretized with linear tetrahedral elements, the average mesh size was 2 mm. Aluminum alloy has been considered as reference material for both components"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001169_1.a32416-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001169_1.a32416-Figure3-1.png",
        "caption": "Fig. 3 Overview of wrapping fold membrane.",
        "texts": [
            " To examine the local buckling on the creased membrane, these research works have to be treated simultaneously. The objective of this paper is to identify the mechanics of the local buckling based on the deformed shape of the creased membrane. To realize the objective, the deformed shape of the creased membrane obtained by our crease model [10] is introduced into the local buckling analysis based on Aksel\u2019rad\u2019s buckling analysis [8,9]. We focus on the z-fold type fold pattern as a simple model of the wrapping fold. An overview of the z-fold type of wrapping fold membrane is indicated in Fig. 3. We make the following three assumptions for the wrapping fold membrane. 1) The cross section of the wrapped membrane is a repeating structure, wherewe assume that many layers of the folded membrane are wrapped around the center hub. Based on the assumption, one wavelength of the repeating structure indicated by the selected area in Fig. 3 is treated. 2) The shape of the center hub is a cylinder, where themembrane is wrapped around the continuous surface. When a polygonal center hub is used, as the curvature of the surface is concentrated at the corner, the behavior of the local buckling is affected by the corner. 3) The tensile force is applied to the membrane so as to wrap the membrane around the center hub. Based on these assumptions, at first, finite-element simulations are demonstrated to examine the mechanical and geometrical properties of local buckling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002099_red-uas.2015.7441020-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002099_red-uas.2015.7441020-Figure1-1.png",
        "caption": "Fig. 1: Convertible UAV.",
        "texts": [
            " The moments generated by the actuators are described by \u0393a = \u00d7 Fact (6) = \u239b \u239d efr e(fe1 + fe2) a(fe2 \u2212 fe1) \u239e \u23a0 (7) The propulsion thrust force Tc = T1 + T2, for T1 = \u03baw2 r1 is the thrust of the motor up and T2 = \u03baw2 r2 for the motor down, is generated by the coaxial rotor, for an aerodynamic constant \u03ba > 0. Then, the vehicle thrust, in B, is given by Fp = Tce3 by assuming that the performance of the upper rotor is not influenced by the lower the rotor, that the rotor planes are sufficiently close together and that each rotor provides an equal fraction of the total system. The motion of the control surfaces is produced by the forces fr, fe1 and fe2, where a, e denoting the distance from the center of mass to the point of generation of those forces, see Fig. 1. Thereby, the vector of the control inputs is \u0393a = (\u03c4\u03c6, \u03c4\u03b8, \u03c4\u03c8) T is given by \u03c4\u03c6 = efr, \u03c4\u03b8 = e(fe1 + fe2) and \u03c4\u03c8 = a(fe2 \u2212 fe1), That is, the 4 independent control inputs, U \u2208 R 4 in matrix form, provides the following well-posed actuation map U = Aaf (8) U = [ T \u0393a ] = \u23a1 \u23a2\u23a2\u23a3 1 0 0 0 0 e 0 0 0 0 e e 0 0 a \u2212 a \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a2\u23a2\u23a3 Tc fr fe1 fe2 \u23a4 \u23a5\u23a5\u23a6 being the map Aa \u2208 R 4\u00d74 invertible with T \u2208 R and \u0393a \u2208 R 3\u00d71 . 2) Due to wind, reaction and gyroscopic effects: The aerodynamic forces of the vehicle, referred to the aerodynamic frame W : B \u2192 W , are generated during the flight, and depend on the wind velocity vector are given by Fa = WT \u239b \u239d Lw Yw Dw , \u239e \u23a0W = \u239b \u239d c\u03b1c\u03b2 s\u03b2 s\u03b1c\u03b2 \u2212c\u03b1s\u03b2 c\u03b2 \u2212s\u03b1s\u03b2 \u2212s\u03b1 0 c\u03b1 \u239e \u23a0 (9) where the notation for cos(\u2217) = c\u2217 and sin(\u2217) = s\u2217 is used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002702_978-3-319-33924-5_13-Figure13.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002702_978-3-319-33924-5_13-Figure13.7-1.png",
        "caption": "Fig. 13.7 Simulation of a neurulation-like process in MGS: from the left to the right, a sheet of epithelial cells is curving until the hems sew together to form a tube",
        "texts": [
            " One of the main difficulties raised by the modeling of these systems is the handling of their dynamical spatial organization: they are examples of dynamical systems with a dynamical structure. In [43] a model of the shape transformation of an epithelial sheet requiring the coupling of a mechanical model with an operation of topological surgery has been considered. This model represents a first step towards the declarative modeling of neurulation. Neurulation is the topological modification of the back region of the embryo when the neural plate folds; then, this folding curves the neural plate until the two borders touch each other and make the plate becomes a neural tube (see Fig. 13.7). Understanding the growth of the shoot apical meristem at a cellular level is a fundamental problem in botany. The protein PIN1 has been recognized to play an important role in facilitating the transport of auxin. Auxin maxima give the localization of organ formation. In 2006, Barbier de Reuille et al. investigated in [4] a computational model to study auxin distribution and its relation to organ formation. The model has been implemented in the MGS language using a Delaunay topological collection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000409_978-3-642-28572-1_38-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000409_978-3-642-28572-1_38-Figure4-1.png",
        "caption": "Fig. 4 Experimental robotic manipulator. The seals have been removed to show the joints.",
        "texts": [
            " A description of a field robotic systems is beyond the scope of this paper. To confirm that unmodeled effects do not invalidate the algorithms described above, an experimental manipulator was designed and fabricated [4]. The manipulator represents the size and kinematic configuration of a well junction field system, given the constraints of the laboratory. For simplicity, only the 3 DOF arm has been implemented, replacing the first prismatic joint with a mounting ring that can be fixed at different heights. The manipulator links have lengths of 8.0 in and 6.0 in (Figure 4). Each joint assembly consists of a motor, gear train, encoder, and associated support bearings. Brushed DC motors are used. The experimental system is sealed to permit it to operate in viscous fluids such as seawater, oil, or drilling mud. The control of the manipulator and the mapping and search algorithms are implemented in C, and run on remote computers using tethers. This is the practice in the oil exploration and drilling industry where wireless communications are not considered practical because of field constraints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003559_s1474-6670(17)70080-8-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003559_s1474-6670(17)70080-8-Figure4-1.png",
        "caption": "Figure 4. (a) Projection of the phase trajectory in the El> E2 plane; (b) Ek) diagram for y \"* 0; (c) E3(T) diagram for y = 0",
        "texts": [
            " There are different possibilities; here the folIowing new phase variables are chosen El = ye' + e\"; E2=ye + e' and E3 = e\" + 2De' + e (6) In the case of no zeros (1X2 = 1X1 = 0; CY-o = 1) these variables satisfy the following system of differential equations dEl dE2 _ E dT = - 2DEl - E2 - Nsgn F; dT - 1 and (7) - yE3 - NsgnF The two first-order differential equations can be replaced by one second-order equation (8) Therefore, the projections into the El' E2 plane (with inclined axes, see reference 8, p. 25) of the trajectory portions between two switchings are logarithmic spirals around the points El = 0, E2 = T N (see Figure 4). The coordinate E3 between two switchings is given by and N E3 = Ce-)'T + - sgn F for y cj= 0 y E3 = - (N sgn F)T + Cl for y = 0 (9) Since the time between two points of the trajectory can be measured in the El' E2 plane [see Figure 4 (a)] the coordinate E3 (normal to the El' E2 plane) can be plotted easily. A con venient arrangement for visualizing the trajectory of a motion is given in reference 9, F{liure 4, p. 37. The coordinates El' E2 and E3 can still be used, when the transfer function has zeros. Equation 2 for a zero-seeking system has on the right-hand side - [1X20\" + ~lb' + :1-015] . Between t This work is supported by the Airforce Office of Scientific Research of the Air Research and Development Command under Contract AF49(638)-513"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003446_aucc.2016.7868196-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003446_aucc.2016.7868196-Figure1-1.png",
        "caption": "Fig. 1. Body fixed principal axes and rotations.",
        "texts": [
            " 978-1-922107-90-9 243 \u00a9 2016 Engineers Australia The paper is organized as follows; a first section introduces the mechanical model of a quadrotor UAV, next the typical sensor equipment is introduced after which a linear model for control synthesis is developed. The linear models of the two different modes are then combined into an RSS model and the MSWS criterion is adopted to the model. Results for LQG design are presented followed by conclusive remarks with suggestions for future directions of related research. The kinematic state of a quadrotor UAV is appropriately captured with angular velocities in body fixed x, y, z (roll, pitch and yaw) directions as shown in figure 1 where \u03c9 = [\u03c9r, \u03c9p, \u03c9y]T denote angular velocity in a body fixed frame. Dynamics are given according to simplified Euler mechanics, i.e. J \u03c9\u0307 = \u03c4 (1) with J = diag(Jp, Jr, Jy) a (diagonal) inertia matrix and \u03c4 = [\u03c4p, \u03c4r, \u03c4y]T the corresponding torque. Note that symmetry assumptions have removed all inertia cross terms, and rotational speed is assumed so low, that any Coriolis effects can be neglected. From angular velocities in global directions g\u03c9 the momentary local to global rotation matrix R = R(t) evolves according to R\u0307 = R \u00d7 g\u03c9 where \u00d7 denotes column-wise vector cross product"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002948_gt2016-57458-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002948_gt2016-57458-Figure6-1.png",
        "caption": "Figure 6 Identification of overhangs at DFP (critical overhands in red \u2013 left: towards cooling plenum; right: towards combustion chamber)",
        "texts": [
            "org/about-asme/terms-of-use Further investigation of the necessary design changes confirmed that the identified overhangs were mainly related to the near-wall cooling channels. It was decided that the impact of these small overhang regions on the overall SLM production is limited and no significant impact on SLM process should be observed. The risk of poor surface quality was already covered with the mentioned increase of the near-wall cooling channel cross section. An overview of the identified overhang regions is presented in Figure 6. As a rule of thumb, the overhang angle should be lower than a certain threshold level (in the range of 40\u00b0 to 50\u00b0) with respect to the substrate plate. The angle depends also on surface quality requirements and other quality requirements. For the SLM version of the damper front panel, the overhang angle for dampers walls was already mitigated to acceptable values due by tilting the panel. However, it was not enough to secure correct manufacturing of part. For the final solution a draft application on the dampers walls in local risk areas was implemented"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003423_icpeices.2016.7853577-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003423_icpeices.2016.7853577-Figure2-1.png",
        "caption": "Fig. 2: Mobile Inverted Pendulum",
        "texts": [
            " It was found that the transfer function of the controlled (MIP) has zeros in the c10sed right half plane (RHP), which come from the c10sed RHP poles appearing in a non-contact loop, the control of such system with FOPID gives better performance [19]. Differential evolution technique for tuning FOPID showed better resuIts than the PSO tuned FOPID [20]. [2] 11. DYNAMIC MODELLlNG The two-wheeled robot consists of a long body with two wheels mounted at the end. For the simplicity of modelling, the two wheels are treated as a unit, and it is assumed that the robot travels only in a straight line. Vnder these conditions the model can be obtained as shown in Fig. 2. The following equations of motion can be obtained as in eq. 1 and 2. 2 .. 2 \" . ' 2 T = (mI + m2 ) 1' BI + (m2L + 12) B2 - m2rLsmB2B2 ( I ) The angle of tiIt from the vertical to the upper body is sufficiently small to allow Iinearization of the system, hence we have (sin (82) ;:::: 82), also the angular velocity of the tiIt of the upper body is sufficiently smalI, such that the centrifugal force may be neglected resuIting in (8 2 )2. The linearized model based on the conditions resuIt in the following equations of motion as given in eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003518_j.engfracmech.2015.05.004-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003518_j.engfracmech.2015.05.004-Figure3-1.png",
        "caption": "Fig. 3. (a) Geometry of a standard freight car wheel with 25-tonnes axle load. The region at the centre of the wheel tread in which the crack is introduced is highlighted. Detailed views of (b) thermal (radial) and (c) RCF (inclined) cracks.",
        "texts": [
            " The following section presents a mesh sensitivity analysis for a wheel with a radial crack and a wheel with a 50 -inclined crack. Both cracks are 1 mm deep (within discretisation tolerances). The wheel sector is subjected to three load traversals. Tractive rolling under full slip with f \u00bc 0:3 is considered. Results are evaluated in the form of CTD magnitudes at P1 in Fig. 5. As a characteristic mesh size, the minimum and maximum element length in the Y 0-direction for elements forming the crack are given in Table 3. In addition, the number of elements and nodes in the region surrounding the crack, see Fig. 3(b), are presented. Mesh sensitivity results with respect to CTDs are presented in Fig. 8. A rather good convergence of CTD magnitudes is achieved. Note that CTD magnitudes also depend on the position of evaluation: due to the mesh discretisation, it is not possible to choose exactly the same position for the different meshes. As a consequence, results for mesh 1 and mesh 3 (where evaluation points are closer) seem to match better than results for mesh 2, which is a finer mesh than mesh 1. It can be concluded that mesh 1 is sufficiently fine and it will be employed in subsequent analyses. It is also seen in Fig. 8 that the dominant displacement mode is shear (mode II). The crack tip is mainly closed as seen in Fig. 8(a). Note that some crack face penetration will occur but that penetration depths are confined to around 1 lm. An additional mesh sensitivity study was carried out for a 1 mm deep 50 -inclined crack. The characteristic mesh size of the crack face region is as previously. The number of elements and nodes in the vicinity of the crack (see Fig. 3(b)), are given in Table 4. Mesh sensitivity results with respect to CTD magnitudes are presented in Fig. 9. A good convergence is achieved. Also for this crack, shear displacement dominates and the crack tip is mainly closed during the rolling cycles as shown in Fig. 9(a). The excessive plastic deformation at the surface in wheel\u2013rail contact limits the use of linear elastic fracture mechanics (LEFM), [7,8,16]. Nevertheless, due to the added numerical complexity in considering elastoplastic material behaviour, use of LEFM is common in the literature, see e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002968_978-981-10-1721-6_48-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002968_978-981-10-1721-6_48-Figure1-1.png",
        "caption": "Fig. 1 Body diagram for the quadrotor and the two reference frames [2]",
        "texts": [
            " The thrust force for each electric motor is driven from the balance of energy Pair = PMech = \ud835\udf02Pelec = \ud835\udf02VinI = \ud835\udf02 Kv Ki \ud835\udf0fm\ud835\udf14m (2) where \ud835\udf02 is the efficiency of the power generation in each electric motor and Pair is the power induced to the air by the electric motor. Based on the momentum theory, we have [2] Pair = \u221a T2 2\ud835\udf0cAswept (3a) T = 3 \u221a (\ud835\udf02 Kv Ki \ud835\udf0fm\ud835\udf14m)2 \u00d7 2\ud835\udf0cAswept (3b) (3) where Aswept is the total area swept by each rotor blade and T is the thrust force. A schematic for body diagram of the quadrotor and the corresponding reference frames are shown in Fig. 1. There are two frames required for defining the quadrotor dynamics. The inertia frame (subscript I) is fixed in earth and the body frame (subscript B) which is fixed to the quadrotor body is rotated by (\ud835\udf03, \ud835\udf19, \ud835\udf13) with respect to the inertia frame. These angles are roll, pitch and yaw which are rotations about x, y and z axes respectively. Here, the rotation matrix is R = \u23a1\u23a2\u23a2\u23a3 C\ud835\udf13.C\ud835\udf19 C\ud835\udf13.S\ud835\udf19.S\ud835\udf03 + S\ud835\udf13.C\ud835\udf03 \u2212C\ud835\udf13.S\ud835\udf19.C\ud835\udf03 + S\ud835\udf13.S\ud835\udf03 \u2212S\ud835\udf13.C\ud835\udf19 \u2212S\ud835\udf13.S\ud835\udf19.S\ud835\udf03 + C\ud835\udf13.C\ud835\udf03 S\ud835\udf13.S\ud835\udf19.C\ud835\udf03 + C\ud835\udf13.S\ud835\udf03 S\ud835\udf19 \u2212C\ud835\udf19.S\ud835\udf03 C\ud835\udf19.C\ud835\udf03 \u23a4\u23a5\u23a5\u23a6 (4) where C\ud835\udf13,C\ud835\udf19,C\ud835\udf03, S\ud835\udf13, S\ud835\udf19 and S\ud835\udf03 denote abbreviated trigonometric functions for cos(\u22c5), sin(\u22c5) respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001699_155892501400900213-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001699_155892501400900213-Figure2-1.png",
        "caption": "FIGURE 2. Forces acting on yarns at crossover points.",
        "texts": [
            " A direct consequence of increase in NL would be decrease in pL, the distance between the load bearing yarns. Olofsson [3] considered the yarn shape as an elastica and proposed that the yarn is bent in shape in a woven fabric by forces acting at the intersections. His work was extended by Grosberg [21] who gave an Eq. 13 for calculating the magnitude of force V required to form the cloth at each intersection. 21 sin 8 vp m \u03b8= \u00d7 (13) where m is the bending modulus of the yarn, \u03b8 is weave angle and p is the inter yarn spacing in the crossing direction. Figure 2 represents the forces acting on the axis of an interlacing yarn in a fabric. Reduction in pL due to increase in NL would result in increase in the binding force of transverse yarns on the load bearing yarn at each interlacement point as shown in Eq. (14), with the subscripts L and T representing the load bearing and transverse directions respectively. 2 8 sin T T T L mv p \u03b8\u00d7 \u00d7 = (14) Since an increase in NL makes the transverse yarns more crimped, they grip a larger section of the surface of load bearing yarn with higher forces at the interlacement point"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002247_s12204-014-1477-7-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002247_s12204-014-1477-7-Figure2-1.png",
        "caption": "Fig. 2 Schematic diagram of the robot arm structure",
        "texts": [
            " It has integrated mobility functions such as walking, climbing wall, crossing obstacles and welding. The robot system consists of a vehicle at the bottom and a manipulator arm mounted on the vehicle, and the welding torch is clamped onto Received date: 2013-06-03 Foundation item: the National High Technology Re- search and Development Program (863) of China (No. 2009AAA042221) and the Fund of Shanghai Sciences & Technology Committee (No. 11111100302) \u2217E-mail: taurus509@163.com the manipulator. The manipulator arm has five degrees of freedom (DOFs). Figure 2 shows the schematic diagram of the robot arm structure, where ai represents the distance of two joint axes along their common perpendicular, and \u03b8i is the angle between two common perpendiculars in the plane perpendicular to the axis. The DenavitHartenberg (D-H) parameters are depicted in Table 1, where \u03b1i is the angle between two joint axes in the plane perpendicular to ai, di is the distance between two common perpendiculars along the axis direction. The connecting rod of the coordinates is constructed to establish the mathematical model in accordance with the D-H rules, and then the robot kinematic equation can be derived"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002424_s1068798x16040109-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002424_s1068798x16040109-Figure5-1.png",
        "caption": "Fig. 5. Crankshaft measurement (a), instrument for shaft measurement by a tactile (contact) method (b); instrument for shaft measurement with an optical measurement system (c); and optical system for the measurement of small parts (d).",
        "texts": [
            " In running the program, prompts to the operator are shown on the screen. As the parame ters are measured, the results and their relationship to the tolerances will be displayed on the screen (Fig. 4). After running the program, all the measurement results may be printed out as a report or stored in a file for further statistical analysis. The MarShaft instruments are universal. At the same time, there are regions where the use of special ized instruments is expedient \u2014 for example, the measurement of crankshafts (Fig. 5a). The operating principles of the instrument remain unchanged, but other functions are added, such as the measurement of crankpin parameters (dynamic measurement of the dynamic). Another specialized application is the mea surement of distributor shafts with additional func tions (for example, measurement of the cam path). MarShaft MAN instruments are supplied in two versions: for the measurement of diameters up to 120 and 220 mm. The height of the centers is 150 mm. In other words, a part of diameter 300 mm extends beyond the frame",
            " Instruments capable of measuring the following lengths are available: 400, 800, 1600, 2000, and 2400 mm. AUTOMATIC INSTRUMENTS Mahr produces two sets of instruments for auto matic measurements. The first group consists of MarShaft CNC instru ments for shaft measurement by a tactile (contact) method. These instruments are analogous in composi tion to the manual instruments (with a horizontal con figuration). However, all the motions of the measuring heads and rotations of the part are controlled automat ically (Fig. 5b). Essentially, this is a numerically con trolled measuring system with a large number of dis placement axes. The measuring process is entirely analogous to that for manual instruments. The measuring time is less, thanks to optimization of the trajectories. The preci sion of the MarShaft CNC instrument in diameter measurement is the same as for manual instruments. The precision in length measurements is slightly higher for the MarShaft CNC instrument than for the MarShaft MAN instrument. These instruments are unable to measure radii and tapers or thread pitches. The parameters in the first three rows of Fig. 2 may be measured. The second group of automatic instruments has an optical measurement system. These instruments have a vertical configuration. The shaft is mounted on a spindle in the lower part of the measurement zone, and the rear headstock is moved downward (Fig. 5c). Two groups of Mahr instruments may be distin guished in terms of the size of the measured part. 316 RUSSIAN ENGINEERING RESEARCH Vol. 36 No. 4 2016 LOKTEV et al. (1) MarShaft SCOPE instruments permit the mea surement of large parts: diameter up to 80 or 120 mm; length up to 350, 750, or 1000 mm. The large spindle diameter prevents losses of preci sion due to the mass of the part (up to 30 kg) or impacts when the part is introduced. Various attach ments may be mounted on the spindle to hold the part: direct and inverse centers, three and six jaw chucks, and spring chucks",
            " To eliminate the contour recognition cycle and the specification of the system dimensions, the system offers the option of adding a module to import a draw RUSSIAN ENGINEERING RESEARCH Vol. 36 No. 4 2016 MEASURING INSTRUMENTS FOR SHAFTS 317 ing of the part in DXF format. The optical system is equipped with a special filter eliminating the influence of dirt at the surface of the part on the measurement precision. In measuring the outline, the dirt particles are filtered from the image in analysis. (2) For the measurement of small parts, Mahr pro duces very compact instruments capable of measuring parts no longer than 200 mm with a diameter no greater than 25 mm (Fig. 5d). All the instruments for shaft measurement may be used directly in production and do not require special buildings or auxiliary equip ment. CONCLUSIONS (1) The instruments for shaft measurement described in the present work must compete in the marketplace with manual measuring instruments, on the one hand, and coordinate measuring machines, on the other. The instruments for shaft measurement are more efficient than manual instruments, since they are able to measure a very large number of parameters in a sin gle setup of the part or even in a single scan"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.16-1.png",
        "caption": "Fig. 3.16 Double wishbone axle of the Audi R8 (with kind approval of the Audi AG)",
        "texts": [
            " bodies joints with each having DoF DoF Fig. 3.15 Planar kinematic chain with closed loops For the kinematic analysis of complex multibody systems with kinematic loops, there exist various approaches for stating the equations of motion. Regarding an automatic assembly of multibody systems, there are mainly three basic approaches which will be demonstrated in the following. They are used as an example for the double wishbone steering axle which is based on the double wishbone wheel suspension of the previous section (Fig. 3.16). It is necessary to have a deep understanding of the topological structure of this suspension (like many other complex multibody systems), as the technical figure alone does not suffice in achieving a well-rounded comprehension of the structure. It can be explained as follows: The system is composed of two coupled kinematic loops and, altogether, has two degrees of freedom \u2013 one for the spring deflection in the suspension mode and the other one for the steering motion (compare Sect. 3.3.2). In order to properly survey the structure of a system with kinematic loops, a graphical representation using the symbols shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001996_sii.2015.7404997-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001996_sii.2015.7404997-Figure1-1.png",
        "caption": "Fig. 1. n-link manipulator whose hand position is constraint by non elastic environment, which is a floor in this figure",
        "texts": [
            " And it has a characteristic different from the conventional method as a force acting on the contact portion from the external environment, not only a constraint force is discussed conventionally, a frictional force exerting contacting point is also formulated. Here, we consider the inverse dynamics solution of con- strained motion of a tip link of straight chain link manip- ulator which constituted by undeformed rigid links while it is contacting the undeformed environment. Considering the manipulator with n rigid links shown in Fig.1, which has a straight chain link structure and n degrees of freedom, and affected by friction force ft and constraint force fn exerted to hand from the ground. We will derive the equation of motion based on a coordinate system \u03a3i fixed to the link i. \u03a30 is a work coordinate system fixed to the floor. The constraint condition can be defined as Eq.(1) when the hand is restrained to a restraint surface, and the r(q) is the position vector of the hand, and q is joint angle vector. C(r(q)) = 0 (1) Here, we can assume that C(r(q)) is differentiable respect- ing to r, q and time t. In Fig.1, which is depicted on 978-1-4673-7242-8/15/$31.00 \u00a92015 IEEE 313 the assumption that the robot is in contact with a floor environment, but the following discussions are not limited to discussions about the floor constraint. First, as a forward dynamics computation of the NewtonEuler method, we can calculate the joint angular velocity i\u03c9i of link i toward the tip link from the root link, the joint angular acceleration i\u03c9\u0307i, the translational acceleration at the origin of \u03a3i, ip\u0308i, and the translational acceleration in the center of mass by the following equation",
            "i) +is\u0302i \u00d7 (mi is\u0308i) + ip\u0302i+1 \u00d7 (iRi+1 i+1fi+1) (8) The ifi, ini in \u03a3i show the force and moment exerted on link i from link (i + 1). And iIi denotes the inertia matrix of the center of gravity of link i. Because n+1fn+1 that is a force transmitting to top link from the floor will be the reaction force of constraint force, we can calculate it as shown in Eq.(6). About constraint motion, we discuss motions with statical and kinetic friction based on friction fundamentals, i.e. (i)The direction of fn and the friction force ft exerted on external contact portion are orthogonal as shown in Fig.1, (ii)ft is determined in proportion to fn: ft = Kfn (K is the coefficient of friction force : 0 < K < 1). The constraint force fn can be determined by the method described in the next chapter. Equation of motion of all links can be obtained by repeating the Newton and Euler\u2019s equation in Eq.(7) and (8) from hand to root link. Giving the \u03a3i to all joints that have rotation axes about the izi-axis, the relationship between ini and joint driving force \u03c4i can be calculated as follows. \u03c4i = izT i ini + Diq\u0307i (9) Here, Di represents the viscous friction coefficient of joint i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure1.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure1.3-1.png",
        "caption": "Fig. 1.3 Porsche 911 a geometry and b finite element mesh",
        "texts": [],
        "surrounding_texts": [
            "1 Introduction to the Finite Element Method 3\nL{y(x)} = b (x \u2208 \u03a9) (1.1)\nand by the conditions which are prescribed on the boundary \u0393 . The differential equation is also called the strong form or the original statement of the problem. The expression \u2018strong form\u2019 comes from the fact that the differential equation describes exactly each point x in the domain of the problem. The operatorL{. . .} in Eq. (1.1) is an arbitrary differential operator which can take, for example, the following forms:",
            "4 1 Introduction to the Finite Element Method\nL{. . .} = d2\ndx2 {. . .}, (1.2)\nL{. . .} = d4\ndx4 {. . .}, (1.3)\nL{. . .} = d4 dx4 {. . .} + d dx {. . .} + {. . .}. (1.4)",
            "Furthermore, variable b in Eq. (1.1) is a given function, and in the case of b = 0, the equation reduces to the homogeneous differential equation: L{y(x)} = 0. More specific expressions of Eqs. (1.3) and (1.4) can take the following form:\na d2y(x)\ndx2 = b, (1.5)\na d4y(x)\ndx4 = b, (1.6)\nand will be used to describe the behavior of rods and beams in the following sections.\n1.1 Example: Draining of a water tank\nGiven is a cylindrical water tank1 (diameter D) which has at x = 0 a small hole of diameter d, see Fig. 1.6a. The initial water-level is h(t = t0) = h0 and the drain velocity is denoted by v(t). Derive the differential equation which describes the draining of the water tank as a function of time. Assume that the water density is constant in the entire tank.\n1This example is adopted from [16]."
        ]
    },
    {
        "image_filename": "designv11_64_0000910_amm.564.412-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000910_amm.564.412-Figure7-1.png",
        "caption": "Fig. 7. The composite specimens before and after failure using extensometer in tensile test",
        "texts": [
            " shows tensile modulus of the composites as a function of the yarn angle. Tensile modulus graph gives the highest value of 10.61 GPa for 0 o loading direction and the tensile modulus for 45 o orientation of loading direction is 5 GPa, followed by the modulus for 90 o of fibre loading direction of 1.2 GPa; which is the lowest. The comparison of strain at different directions of kenaf yarn fibre UP composites (Fig. 3 and 6) reveals that the 45 o gives the best results of stress vesus strain performance. Fig. 7 shows the composite specimens at 45\u00ba orientation of before and after failure. Crack direction is the same with orientation fibres in composites. The SEM micrographs in Fig. 8 represent the fracture surfaces of kenaf yarn fibre composites after the composites were subjected to tensile testing. By examining the SEM micrographs of kenaf yarn fibre UP composites of 0\u00b0 fibre orientation in Fig. 8(a), it can be observed that extensive fibre pull-out was observed from the fracture surface of kenaf yarn fibre UP composites at 0\u00b0 fibre orientation in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001477_s00170-015-7015-4-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001477_s00170-015-7015-4-Figure10-1.png",
        "caption": "Fig. 10 Effect of the maximum transient cutting force on workpiece",
        "texts": [
            " The experiments focused on three areas: the effect of transient cutting force on the structural stress in the first cutting step, the residual stress in the coating, and the bond strength calculation model based on the cutting parameters for the second cutting step. Figure 9a, b, respectively, show the time-domain cutting force in the first and second cutting steps. There was a transient change in the cutting process due to the coating\u2019s coarse surface which indicated that the maximum cutting force was about 160 N with a fade time of 160 ms, so there was effect of forced vibration factor due to the first transient cutting force on the coating cutting process. Then, the transient structural stress in the workpiece is studied (see Fig. 10): Fig. 10a shows the FEM-based transient calculation model of the spindle-workpiece vibration system based on Fig. 2, and it includes bearing springs, spindle structures, and the workpiece. The time-domain cutting force in Fig. 9 was located within one circle of the cutting feed, therefore the transient cutting force was located on this circumference, and the direction of the cutting force followed the orientation of the spring. The results are shown in Fig. 10b, c, which show structural transient stresses in the whole system and cross-sectional stresses at maximum stress. The stress data showed that the maximum transient stress was 7.27 MPa at 80 ms under the maximum force of 160 N. The zone of influence was about 10-mm thick at 4.68 to 7.27 MPa, so the bond strength of the coating was influenced by the transient cutting force as suggested. To study the effect of cutting residual stress on Fe-Al-based coatings, the measured residual stress was assessed using an X-350A device: the data from layers at 20-\u03bcm depth increments and the experimental parameters are summarized in Table 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002778_j.proeng.2016.07.095-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002778_j.proeng.2016.07.095-Figure1-1.png",
        "caption": "Fig. 1.Gearandwheelblanks: (a) sketch of gear section; (b) sketch of key way; (c) gear blank; (d) sketch of wheel section; (e) sketch of splined bore; (f) wheel blank",
        "texts": [
            " The primary focus was placed on creation of a geometrically accurate involute gear tooth contour in the normal to the tooth curves plane. Tooth rims were formed through distributing teeth along wheel disks by Circular Sketch Pattern. For modeling we choose gear parameters, given in the source [1], except the face width, which we increase for more accurate determination of a contact pattern. The given and calculated parameters [9, 10-12] are given in the table. At first, we develop wheel blanks. In the plane of the front view we make sketches of an axial section of the gear (Fig. 1, a) and the wheel (Fig. 1, d). We place an initial point on the rotation axis so that the plane of the right view would be a symmetry plane. Having applied the command Revolved Boss/Base, we form gear wheel blanks. In the transverse plane of hubs we draw sketches of a key way (Fig. 1, b) and a splined bore (Fig. 1, d). By the command Extruded Cut we make a bore in hubs, rounding and chamfers (Fig. 1, c, f), by doing so we form gear blanks. 3. Construction of tooth rims of a gear and a wheel The normal module mn is standardized. It\u2019s initial in case of geometrical and strength calculations of the helical gear. Therefore, we make a contour of the helical gear tooth in the Plane which is perpendicular to the tooth curve. The construction of the tooth rim we start with a helix g, which specifies a direction of the tooth curve. We carry out the following actions in order a middle point of the helix could coincide with the pitch point P (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003350_cefc.2016.7815913-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003350_cefc.2016.7815913-Figure3-1.png",
        "caption": "Fig. 3. The magnet flux density distribution. (a) Regular PMV machine. (b) Proposed PMV machine.",
        "texts": [
            " It is seen that the consequent pole in the rotor is employed, and the Halbach array PMs in the stator opening is used to guide the flux through airgap into the stator as shown in Fig. 2. As shown in Figs. 1 and 2, only the S-pole magnets produce the main flux, and N-pole magnets produce the leakage flux to reduce main flux. In the proposed PMV machine, almost all the magnets produce the main flux. Hence, the significant improvement on the magnet flux can be obtained with similar magnet usage as shown in Fig 3. The back-EMF and torque versus current curves of the proposed and the regular PMV machines are shown in Fig. 4. It can be seen that the proposed PMV machine produces 85% larger back-EMF than that of the regular PMV machine.When the RMS value of line current of two machines are both 7.25 A, proposed PMV machine produces 49% larger torque than that of the regular PMV machine. And with current decrease, the increase proportion of torque by proposed PMV machine becomes larger. Besides, when the torque of two machines are both 20"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001630_j.proeng.2014.03.133-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001630_j.proeng.2014.03.133-Figure7-1.png",
        "caption": "Fig. 7. Scheme of vibration protection of object. Fig. 8. Analytical model of vibrating object.",
        "texts": [
            " Stiffness curve may be approximated by linear dependence on displacement with formula: 671081210310 .z.zcc)z(cc . (13) For the dependence \u201cforce-displacement\u201d in accordance with linear theory c=108.45 kN/mm. on axial force: experimental data, on vertical displacement: experimental data, - in accordance with linear theory. - in accordance with approximating curve. Laminated elastomeric vibroisolator with discussed above characteristics is placed between the object to be protected and a vibrating base (Fig. 7). The lower plate of the vibroisolator is subjected to kinematic excitation. In this paper the periodic excitation \u03be(t) = \u0394sin(\u03c9t), is taken for numerical solution. It is assumed that the external excitation is independent of motion of the system to which it is applied. Determination the law of motion of the upper plate, on which the protected object is located, is the important problem of passive systems. In this case excited vibration amplitudes of plate depend on the excitation frequency and on the possibility of resonance phenomenon occurrence in an oscillating system \"protected objet - vibroisolator\""
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003111_icemi.2015.7494242-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003111_icemi.2015.7494242-Figure6-1.png",
        "caption": "Fig 6. The human machine interface of Condition",
        "texts": [
            " Displacement sens ors measure dynamic displacement and are used to collect signaIs from low speed shaft and gear. The CMS uses sensors to automatically collect data according to a user-defined schedule. After data collection, the measurements are transferred via a TCP /IP based Ethernet network to a desktop PC or server running the condition monitoring software. RJ-45 sockets on CMS device can be used link the devices to the network. CMS acquires vibration signal online and sends them to condition monitoring software to analyze online or offline. Fig. 6 shows the main panel of the data processing and analyzing. The part l of Fig. 6 demonstrates the 3 dimension diagram of WT gearbox test rig. The part II shows vibration data from CMS in time domain of Fig. 6 (a) or frequency domain of Fig. 6 (b) as requirement. The part III indicates component with fault and fault type according to analysis of vibration data. monitoring software. The purpose of the test rig is to produce fault-like signal in order to diagnose fault in time domain and frequency domain. lnspired by research of Durham University, the low speed end of gearbox has been fitted with experimental unbalance plane to examine low speed shaft mass unbalance [1 Ol Holes at four different radii are drilled at 90\u00b0 intervals around the unbalance plane fitted to low speed shaft of gearbox"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.34-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.34-1.png",
        "caption": "FIGURE 6.34",
        "texts": [
            " There are a number of steering system configurations available for cars and trucks based on linkages and steering gearboxes. The treatment in the following sections is limited to a traditional rack and pinion system. For the simple full vehicle models discussed earlier, such as that modelled with lumped mass suspensions, there are problems when trying to incorporate the steering system. Consider first the arrangement of the steering system on the actual vehicle and the way this can be modelled on the detailed linkage model as shown in Figure 6.34. In this case only the suspension on the right hand side is shown for clarity. The steering column is represented as a part connected to the vehicle body by a revolute joint with its axis aligned along the line of the column. The steering inputs Modelling the steering system. required to manoeuvre the vehicle are applied as motion or torque inputs at this joint. The steering rack part is connected to the vehicle body by a translational joint and connected to the tie rod by a universal joint. The translation of the rack is related to the rotation of the steering column by some kind of coupling statement that defines the ratio; such constructs are common to most general purpose software packages"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002136_0954406216639343-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002136_0954406216639343-Figure1-1.png",
        "caption": "Figure 1. GTS system with deadzone.",
        "texts": [
            " A baseline smooth backstepping controller is designed for the nominal model of the GTS system, and the uniformly ultimate boundedness of the output tracking error is guaranteed by the robustness analysis for the system in the presence of deadzone nonlinearity with parametric uncertainties. Simulation results are presented at the end to show the effectiveness of the proposed controller and the high performance without limit cycles of the closed-loop responses. Problem statement Dynamic model of GTS systems A typical dynamical model for GTS systems (as shown in Figure 1) can be expressed as4 Jm d2 m dt2 \u00fe cm d m dt \u00bc u Dead\u00f0 \u00de Jl d2 l dt2 \u00fe cl d l dt \u00bc N0Dead\u00f0 \u00de ( \u00f01\u00de where Jm, m and cm are the moment of inertia, position and viscous friction coefficient of the driving side, respectively, Jl, l and cl the moment of inertia, position and viscous friction coefficient of the load side, respectively, N0 the reduction ratio, u the torque input, \u00bc m N0 l the relative displacement, and Dead( ) the transmission torque, defined by Dead\u00f0 \u00de \u00bc k\u00f0 r\u00de, if 5 r 0, if 2 \u00f0 l, r\u00de k\u00f0 \u00fe l\u00de, if 4 l 8>< >: \u00f02\u00de with the rigidity coefficient k and the break points l< 0 and r> 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.17-1.png",
        "caption": "Fig. 10.17 Milling schemes for finish milling using either a five-axis ballend mill operation mode (left) or a five-axis flat-end mill operation mode (right)",
        "texts": [
            " Unlike the rough machining, the finish machining step requires high accuracy to fulfill the final product specifications. The corresponding tool path is aimed at removing the remaining material over the expected product surface after the rough machining. In other words, the finish machining performs a few additional tasks like smoothing surfaces and edges, fixing angles, or removing excessive teeth. Taking milling as an example, it is common to apply the machining mode with five axes or with three axes equipped with a ball-end mill (Fig. 10.17) as these configurations can generate curved surfaces, fillet edges, and inclined surfaces without the tooth-alike surface profiles. Once the machining configuration has been set, the tool path can then be computed. For instance, the length of each line of machine tool movement is flexible. Such length can be adjusted based on the required tolerance and the final shape 288 10 Design for Manufacturing features, e.g., inclined angle of side surfaces and the radius of fillet edges, as exemplified in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.19-1.png",
        "caption": "FIGURE 8.19",
        "texts": [
            " The road profile is characterized by the geometric shape of a profile cam, which is assembled to the tire using a cam-follower connection. The quarter suspension consists of major components that essentially define the kinematic and dynamic characteristics of the racecar. These components include upper and lower control arms, Formula SAE racecar designed and built by engineering students: (a) physical racecar and (b) virtual racecar designed in Pro/ENGINEER. Right front quarter of the racecar suspension. upright, rocker, shock, push rod, tie rod, and wheel and tire, as shown in Figure 8.19. The dangling end of the shock, both control arms, rocker, and tie rod are connected to the chassis frame using numerous joints. The chassis frame was assumed fixed and the tire is pushed and pulled by the profile cam (not shown), mimicking the road profile. Two views, shown in Figure 8.19, were created to aid the visualization of the assembled model. The tire of the quarter suspension is in contact with the profile cam that characterizes the road profile. As shown in Figure 8.20(a), the geometry of the cam consists of two circular arcs of radius 7.65 in. (AB and FG), which are concentric with the cam center. Therefore, when the cam rotates, these two circular arcs do not push or pull the tire; the result is two flat segments of the road profile, as shown in Figure 8.20(b). In addition, the circular arc CDE is centered 4 in"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003236_detc2016-59619-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003236_detc2016-59619-Figure1-1.png",
        "caption": "FIGURE 1. N DEGREES OF FREEDOM SERIAL CHAIN",
        "texts": [
            " The kinematic analysis, the stability criteria and the control scheme for the manipulation are provided in Section 3. Simulations and results are shown in Section 4. Classically, the center of mass (COM) of a multi-body system is defined as the weighted average of the positions of each individual body\u2019s COM, x\u0304n = \u2211 n i=1 mic\u0304i \u2211 n i=1 mi where c\u0304i is used to locate the COM of an individual body in the world reference frame, and the mass of body i is given by mi. Consider a kinematic chain (shown as Figure 1) composed of n rigid links Li, connected by revolute joints. For each joint, construct a twist \u03bei which corresponds to the screw motion of the ith joint with all other joint angles fixed at \u03b8 j = 0. Let us consider the simplest condition first. Assume there is only one link L1, as shown in Figure 1.a. The COM of this manipulator can be represented as [ x\u03041 1 ] = e\u03b81\u03be\u03021 [ c\u03041 1 ] Then we connect a new link L2 to L1, shown as in Figure 1.b. Suppose that we fix joint 2 and study the COM of the new system as a function of m2 only. Then, the following equation can be obtained, x\u03042 = 1 \u2211 2 i=1 mi (m1x\u03041 +m2c\u03042), which can be written in a recursive form as, [ x\u03042 1 ] = \u03c11M1 [ x\u03041 1 ] , \u03c11 = m1 m1 +m2 , M1 = [ I3\u00d73 m2 m1 c\u03042 0 m1+m2 m1 ] , Next, when considering the rotation of joint 2, by composition, the COM becomes[ x\u030412 1 ] = e\u03b82\u03be\u03022\u03c11M1e\u03b81\u03be\u03021 [ c\u03041 1 ] . Finally, for a kinematic chain composed of n rigid links Li, we can write the expression of COM as [ x\u0304 1 ] = \u03c11\u03c12 \u00b7 \u00b7 \u00b7\u03c1n\u22121e\u03b8n\u03be\u0302nMn\u22121e\u03b8n\u22121\u03be\u0302n\u22121 \u00b7 \u00b7 \u00b7M1e\u03b81\u03be\u03021 [ c\u03041 1 ] = \u03c11\u03c12 \u00b7 \u00b7 \u00b7\u03c1n\u22121e\u03b8n\u03be\u0302nMn\u22121e\u03b8n\u22121\u03be\u0302n\u22121M\u22121 n\u22121Mn\u22121Mn\u22122e\u03b8n\u22122\u03be\u0302n\u22122M\u22121 n\u22121 M\u22121 n\u22122 \u00b7 \u00b7 \u00b7e \u03b81\u03be\u03021M\u22121 n\u22121 \u00b7 \u00b7 \u00b7M \u22121 1 M1 \u00b7 \u00b7 \u00b7Mn\u22121 [ c\u03041 1 ] = e\u03b8n\u03be\u0302ne\u03b8n\u22121\u03be\u0302 \u2032 n\u22121 \u00b7 \u00b7 \u00b7e\u03b81\u03be\u0302 \u2032 1 \u03c11 \u00b7 \u00b7 \u00b7\u03c1n\u22121M1 \u00b7 \u00b7 \u00b7Mn\u22121 [ c\u03041 1 ] = e\u03b8n\u03be\u0302ne\u03b8n\u22121\u03be\u0302 \u2032 n\u22121 \u00b7 \u00b7 \u00b7e\u03b81\u03be\u0302 \u2032 1 [ x\u03040 1 ] (1) 2 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure15-1.png",
        "caption": "Fig. 15. Girling Sliding Shoe Diagram",
        "texts": [
            " At 250 deg. C. compounds were found which had a high molecular weight, and a free OH group, and those, of course, would constitute excellent lubricants. Mr. G. A. G. FAZEKAS, A.M.I.Mech.E., commented on the author\u2019s statement that when the brakes were released the floating shoe required some kind of frictional device to centre it in the off-position, and irregularities in the rotation of the drum had to be relied upon to knock the shoe clear. Such a condition did not arise with the new Girling design. Fig. 15 showed, in a skeletonized form, but with the main dimensions, a 9-inch by 14-inch brake. The shoe rested at point 0 on the abutnient, and, assuming that in the off-position there was no tendency to slide on application, 0 would be the instantaneous centre of rotation. In that case the heel of the lining cleared the drum well. If the heel was more or less under the abutment, as was the case with some brakes, the position was not so good, and the drum might have to be relied upon to knock the shoe clear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure12-1.png",
        "caption": "Figure 12 Deflection of 3P4 specimen configuration (the unit is in mm)",
        "texts": [
            " four coordinate systems were assigned to four specimen configurations 3P1, 3P2, 3P3, and 3P4. An additional rotation of 45\u00b0 about the z-axis was made to coincide the global coordinate system of ABAQUS with the machine coordinate system. Specimens were prescribed with the same boundary condition with the right and left bottom line fixed. An evenly distributed load of total 200N was applied on all the nodes along the loading line. The deflection contour in direction 3 of 3P4 specimen is shown in Figure 12. As expected, it is observed that highest deflection locates in the middle of the specimen. Table 3 shows deflections of specimens tested under 200N from FEA results. The deflection of the beam is also calculated using the strength of materials approach. The 3-point bending test was simplified to a model of a simply supported beam with a concentrated load at the centre. Since the thickness of the beam is comparable to its width and length, total deflection is due to both bending and shear forces: t b s    (8) where \u03b4t is the total deflection of the beam, \u03b4b is the maximum deflection in the middle of the specimen due to bending, \u03b4s is the maximum deflection due to shearing strain"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.30-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.30-1.png",
        "caption": "Fig. 2.30 Free body diagram of the truss structure: a force boundary condition; b displacement boundary condition",
        "texts": [
            " The three trusses have the same length L , the same Young\u2019s modulus E , and the same cross-sectional area A. The structure is loaded by (a) a horizontal force F at node 2, (b) a prescribed displacement u at node 2. Determine for both cases \u2022 the global system of equations, \u2022 the reduced system of equations, \u2022 all nodal displacements, \u2022 all reaction forces, \u2022 the force in each rod. 2.4 Assembly of Elements to Plane Truss Structures 65 2.5 Solution The free-body diagram and the local coordinate axes of each element are shown in Fig. 2.30. From this figure, the rotational angles from the global to the local coordinate system can be determined and the sine and cosine values calculated as given in Table2.15. (a) Force boundary condition Based on Eq. (2.190) and the values given in Table2.15, the elemental stiffnessmatrices can be calculated as: Element Angle of Rotation sine cosine I 30\u25e6 1 2 \u221a 3 2 II 90\u25e6 1 0 III 330\u25e6 \u22121 2 \u221a 3 2 K e I = E A L\ufe38\ufe37\ufe37\ufe38 kI \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 3 4 \u221a 3 4 \u2212 3 4 \u2212 \u221a 3 4\u221a 3 4 1 4 \u2212 \u221a 3 4 \u2212 1 4 \u2212 3 4 \u2212 \u221a 3 4 3 4 \u221a 3 4 \u2212 \u221a 3 4 \u2212 1 4 \u221a 3 4 1 4 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 , (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001518_2015-01-9087-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001518_2015-01-9087-Figure2-1.png",
        "caption": "Figure 2. Spiral Groove Journal Bearing",
        "texts": [
            " The main objective of this paper is to describe the selection of the robust design parameters for a Spiral Groove Journal Bearing, which not only meet the load carrying capacity requirements, but also have the minimum sensitivity to both internal and external noises such as clearance, misalignment and temperature of the lubricating fluid. Taguchi's cross orthogonal array Design of Experiment (DOE) procedure [16, 17] is used to arrive at a robust solution. A schematic of a Spiral groove journal bearing is shown in Figure 2. The main geometrical parameters for a spiral groove journal bearing are- the diameter and length of the bearing, as well as the width, height and pitch of the groove. For the robust design process, all the operational requirements and design variables were first obtained, and then classified as per their impact on the Journal bearing performance. Figure 3 shows the Parameter Diagram (P-Diagram) for a journal bearing. Three important noise factors - clearance, misalignment and temperature, each with two levels, were selected for the Robustness study"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure1.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure1.10-1.png",
        "caption": "Fig. 1.10 a Active two-pole without a voltage stabilization. b Stabilization of a load voltage",
        "texts": [
            " Then, there is a problem, how reasonably we may express these parameters for the initial regime Y1 L2, subsequent Y 2 L2 and regime changes in the relative form. It is possible to note that regime changes, as the length of segments on all the load straight lines have different lengths for usually used Euclidean geometry. 1.3 Analysis of the Traditional Approach to Normalizing of Regime Parameters for the Voltage Linear Stabilization For the illustration of the assigned task, we consider two simple active two-poles with a load conductivity YL1 in Fig. 1.10. The equation of the load straight line of the first active two-pole in Fig. 1.10a is given by I1 \u00bc \u00f0V0 V1\u00dey01 \u00bc y01V0 y01V1; where the conductivity y01 corresponds to the internal resistance of the voltage source V0 and SC current ISC1 \u00bc y01V0. Then, we use normalized expression (1.2) which contains the two normalized values I1 ISC1 ; V1 V0 : 1.2 Disadvantages of the Well-Known Calculation Methods of Regime \u2026 11 The regimes of two similar circuits with running values of currents I1; ~I1 and voltages V1; ~V1 will be equivalent or equal to each other if the normalized values of currents and voltages (1.3) take place; that is, I1 ISC1 \u00bc ~I1 ~ISC1 ; V1 V0 \u00bc ~V1 ~V0 : The equation of the load straight line of the second active two-pole in Fig. 1.10b is given by I1 \u00bc y0Ny1N y0N \u00fe y1N \u00f0V0 V1\u00de; where conductivity y1N corresponds to the conductivity of voltage regulator. It is possible to carry out the normalization by the SC current ISC1 if there is an access to this source at an experimental investigation. Then, I1 ISC1 \u00bc y1N=y0N 1\u00fe y1N=y0N 1 V1 V0 : \u00f01:19\u00de This expression contains the three normalized values. Therefore, a possible condition of equal regimes corresponds to the equalities I1 ISC1 \u00bc ~I1 ~ISC1 ; V1 V0 \u00bc ~V1 ~V0 ; y1N y0N \u00bc ~y1N ~y0N : \u00f01:20\u00de If regimes differ, how may we express this difference in the convenient view"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure11.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure11.9-1.png",
        "caption": "Fig. 11.9 Different configurations of bioreactors. (a) Incubator with spinner. (b) Incubator with rotary chamber wall. (c) Perfusion chamber. (d) Hollow tubes with culture media perfusion",
        "texts": [
            " Incubator-based bioreactors can regulate nutrient flow, gaseous ions, and signaling molecule levels. Multi-pass filtration seeding can be applied to place cells in three-dimensional scaffolds. Apparently, this approach produces more uniformly distributed cells inside scaffolds than static seeding, i.e., cells which are just pipetted onto the scaffold body inside culture media. Furthermore, the flow transport of oxygen and soluble nutrients can be enhanced through culturing the construct in a stirring spinner flask (Fig. 11.9a). The spinner generates semi-controlled fluid shear which produces detrimental turbulent eddies which mix oxygen and nutrients, thereby reducing the boundary layer at the construct surface from a reduction of the diffusional limits through the generation of a dynamic laminar flow with minimal levels of shear. Other than agitation using a spinner, the chamber can be configured with a rotating wall (Fig. 11.9b) to induce low shear stresses and high mass transfer rate. The rotating wall has an advantage of balancing forces to stimulate \u201czero gravity\u201d conditions, meaning that the cells, nutrients, and biowastes do not accumulate at the chamber base due to gravity. In addition to the incubator chambers, perfusion techniques can be integrated into a bioreactor system. Culture media are continuously perfused through the scaffolds placed inside the perfusion chamber (Fig. 11.9c). The culture media that exited the chamber may then recirculate at the inlet of the chamber, or alternatively, fresh culture media can always be supplied from the inlet. In this case, the continuous flow ensures nutrients and dissolved oxygen are continuously supplied to cells in the scaffold and the biowastes from cells do not accumulate around cells. Besides, the perfusion bioreactors can be inserted with the semipermeable hollow fibers (Fig. 11.9d) which contain both cells and scaffolds with the fiber direction perpendicular to the media flow. This approach maintains the cells in the scaffolds and culture regions in the hollow fibers while also allowing for the required mass 11.5 Implementation of In Vitro Tissue Regeneration 317 transport of oxygen and nutrients to the highly metabolic cells. Furthermore, these bioreactors can also be applied with mechanical conditioning as the scaffold is physically contacted with the fibers. Cyclical mechanical stretch on the fibers can induce direct stresses on the scaffolds and the cells inside"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure5.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure5.3-1.png",
        "caption": "Fig. 5.3 Two dimensional problem: plane strain",
        "texts": [
            "7) 244 5 Plane Elements where D = C\u22121 is the so-called elastic compliancematrix. The general characteristic of plane Hooke\u2019s law in the form of Eqs. (5.5) and (5.6) is that two independent material parameters are used. It should be finally noted that the thickness strain \u03b5z can be obtained based on the two in-plane normal strains \u03b5x and \u03b5y as: \u03b5z = \u2212 \u03bd 1 \u2212 \u03bd \u00b7 ( \u03b5x + \u03b5y ) . (5.8) The last equation canbederived from the tree-dimensional formulation, seeSect. 7.1.2 5.2.2.2 Plane Strain Case The two-dimensional plane strain case (\u03b5z = \u03b5yz = \u03b5xz = 0) shown in Fig. 5.3 is commonly used for the analysis of elongated prismatic bodies of uniform cross section subjected to uniform loading along their longitudinal axis but without any component in direction of the z-axis (e.g. pressure p1 and p2), such as in the case of tunnels, soil slopes, and retaining walls. It should be noted here that the normal thickness strain is zero (\u03b5z = 0) whereas the thickness normal stress is present (\u03c3z = 0). The plane strain Hooke\u2019s law for a linear-elastic isotropic material based on the Young\u2019s modulus E and Poisson\u2019s ratio \u03bd can be written for a constant temperature as \u23a1 \u23a3 \u03c3x \u03c3y \u03c3xy \u23a4 \u23a6 = E (1 + \u03bd)(1 \u2212 2\u03bd) \u23a1 \u23a3 1 \u2212 \u03bd \u03bd 0 \u03bd 1 \u2212 \u03bd 0 0 0 1\u22122\u03bd 2 \u23a4 \u23a6 \u00b7 \u23a1 \u23a3 \u03b5x \u03b5y 2 \u03b5xy \u23a4 \u23a6 , (5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.5-1.png",
        "caption": "FIGURE 8.5",
        "texts": [
            " Multibody kinematic analysis involves formulating equations of motion and solving them for position, velocity, and acceleration of individual bodies in the system in time. Such an analysis is important for general mechanism analysis and design, particularly for workspace analysis and robotics, where position, velocity, and acceleration of the moving components must be known in order to assess the functionality and performance of the mechanical system. One of the most famous mechanisms commonly mentioned in spatial kinematic analysis is the Stewart platform, as shown in Figure 8.5(a). The Stewart platform mechanism, originally suggested by Stewart (1965), is a parallel kinematic structure that can be used as a basis for controlled motion with six degrees of freedom (DOF). The mechanism itself consists of a stationary platform (base platform) and a mobile platform that are connected via six legs (struts) mounted on universal joints. The legs have a built-in mechanism that allows changing the length of each individual leg. The desired position and orientation of the mobile platform are achieved by combining the lengths of the six legs, transforming the six transitional DOF into three positional (displacement vector) DOF and three orientation DOF (angles of rotation of a rigid body in space). The advantage of the Stewart platform is six degrees of freedom and a split-hair accuracy of mobile platform positioning. This makes Stewart platform widely usable in robot-building infrastructures. For example, such a parallel kinematic structure is used in the National Advanced Driving Simulator (NADS) shown in Figure 8.5(b). In this section, instead of analyzing complex mechanisms such as the Stewart platform, which involves complex mathematical formulations, we discuss simple applications that can be solved analytically. One of the simplest and most useful mechanisms is the four-bar linkage, among which the slider-crank mechanism (Figure 8.6) can be seen in many applications such as internal combustion engines and oil-well-drilling equipment. For the internal combustion engine, the mechanism is driven by a firing load that pushes the piston, converting the reciprocal motion into rotational motion at the crank"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003420_ecce.2016.7855092-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003420_ecce.2016.7855092-Figure2-1.png",
        "caption": "Fig. 2. Basic structure of V-shaped IPMSM using concentrated winding structure.",
        "texts": [
            " This paper provides the examination of rotor structures which use permanent magnet torque effectively by suppressing the q-axis magnetic flux and by concentrating the field magnetic flux of permanent magnets (PMs) on the d-axis in order to enhance the efficiency of V-shaped IPMSM using concentrated winding structure at high speed and high torque area. II. BASIC STRUCTURE OF V-SHAPED IPMSM Table I shows specification of the examined V-shaped IPMSM. A 6-pole/9-slot of integral slot, which has small lower harmonic iron loss, is adopted to achieve high efficiency operation at high speed and high torque area [12]- [14]. Also, concentrated winding structure is adopted to cut manufacturing cost and to use limited space effectively. Fig. 2 shows basic structure of V-shaped IPMSM using concentrated winding structure. The stator core and the rotor core are made of silicon steel lamination of 35A210. NMX41SH made by Hitachi Metal, Ltd. is used as Nd-Fe-B magnets for the rotor. III. EXAMINATION OF THE ROTOR STRUCTURES The characteristics of the V-shaped IPMSM using concentrated winding structure were examined at high speed and high torque area by means of employing the JMAGDesigner simulation software. Four rotor structures models which have the same volume and shape of magnets are used to measure the effect of suppressing the q-axis magnetic flux and concentrating the field magnetic flux of PMs on the daxis. Fig. 3(a) shows the rotor structure of model A. Note that the basic structure shown in Fig. 2 is defined as model A. There is minimum flux barrier on the q-axis magnetic flux path in this model. Therefore, this model is the easiest structure for generating reluctance torque because its salient pole ratio is the biggest among other examined structures. Fig. 3(b) shows the rotor structure of model B. There is disproportional gap on q-axis to suppress the q-axis magnetic flux and its harmonic component in this model. Fig. 3(c) shows the rotor structure of model C. There is large flux barrier on the q-axis magnetic flux path to concentrate the field magnetic flux of PMs on d-axis while suppressing the q-axis magnetic flux in this model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.24-1.png",
        "caption": "Fig. 3.24 Free body diagram of the beam structure problem",
        "texts": [
            " The left-hand part of the structure (0 \u2264 X \u2264 L) is loaded by a constant distributed load q. Use two Euler\u2013Bernoulli beam elements of equal length (see Fig. 3.23b) and: \u2022 Assemble the global system of equations without consideration of the boundary conditions at the fixed supports. \u2022 Obtain the reduced system of equations. \u2022 Solve the system of equations for the unknowns at node 2. \u2022 Calculate the reactions at node 1 and 3. 134 3 Euler\u2013Bernoulli Beams and Frames 3.3 Solution The free body diagram is shown in Fig. 3.24. \u2022 Let us look in the following first separately at each element. The stiffness matrix for element I can be written as: Ke I = EIY L3 u1Y \u03d51Z u2Y \u03d52Z\u23a1 \u23a2 \u23a2 \u23a3 12 \u22126L \u221212 \u22126L \u22126L 4L2 6L 2L2 \u221212 6L 12 6L \u22126L 2L2 6L 4L2 \u23a4 \u23a5 \u23a5 \u23a6 u1Z \u03d51Y u2Z \u03d52Y . (3.170) In the same way, the stiffness matrix for element II reads as: Ke II = EIY L3 u2Y \u03d52Z u3Y \u03d53Z\u23a1 \u23a2 \u23a2 \u23a3 12 \u22126L \u221212 \u22126L \u22126L 4L2 6L 2L2 \u221212 6L 12 6L \u22126L 2L2 6L 4L2 \u23a4 \u23a5 \u23a5 \u23a6 u2Z \u03d52Y u3Z \u03d53Y . (3.171) The global system of equation without consideration of the boundary conditions is obtained as: EIY L3 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 12 \u22126L \u221212 \u22126L 0 0 \u22126L 4L2 6L 2L2 0 0 \u221212 6L 12 + 12 6L \u2212 6L \u221212 \u22126L \u22126L 2L2 6L \u2212 6L 4L2 + 4L2 6L 2L2 0 0 \u221212 6L 12 6L 0 0 \u22126L 2L2 6L 4L2 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 u1Z \u03d51Y u2Z \u03d52Y u3Z \u03d53Y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 R1Z \u2212 qL 2 \u2212M1Y + qL2 12 \u2212 qL 2 \u2212 qL2 12 R3Z \u2212M3Y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure5-1.png",
        "caption": "Figure 5 Locking mechanism (enlarged view) (see online version for colours)",
        "texts": [
            " Every passive joint of the manipulator has a locking mechanism, and a hose for the locking mechanisms is installed along the centre of the manipulator. A rail is installed on each manipulator. The two manipulators are connected through the rails and can move along each other via the rails [Figure 2(b)]. As shown in Figure 3, the head link of each manipulator can be turned by pulling either wire. As shown in Figure 4, a piston is attached to the central hose. Each manipulator has its own piston. By pushing on the piston, the hose expands until the two friction materials are engaged. The joint is thus locked (Figure 5). The proposed duplex mechanism enables movement as follows (Figure 6). First, at the start of a curve, one manipulator is moved forward. Next, the direction of the head of the manipulator is changed by pulling a wire. Then, all of the joints of the manipulator are locked by pushing on the piston. Next, the other manipulator is moved forward along the locked manipulator by pushing on its rear end [Figure 6(b)]. Then, the locked manipulator is released, and all of the joints of the other manipulator are locked"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure13.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure13.11-1.png",
        "caption": "Fig. 13.11 Dependence Ri\u00f0V\u00de for a given load power and conformity of different regime parameters",
        "texts": [
            "9 Characteristics of a power-load element with different values of an internal resistance S \u221e \u221e V M M \u2212 M + V 0 V I 0 13.3 Influence of Voltage Source Parameters and Power-Load Element \u2026 371 Therefore, the single-valued mapping of the parabola points onto the axis V takes place. Using expression (13.7), we may constitute the cross-ratio m1 V for the initial point V1; that is, m1 V \u00bc VM V1 V0 1 \u00bc V1 VM V0 VM \u00bc V1 V0=2 V0=2 : \u00f013:25\u00de The points VM \u00bc V0M 2 ; V \u00bc 1 are base ones and V0 is a unit point. The conformity of the points V1, m1 V is shown in Fig. 13.11. Similarly to the above, let us consider the cross-ratio m1 i for the resistance Ri using the conformity of the variables by Fig. 13.11 for the single-valued area. This cross-ratio has the form m1 i \u00bc RiM R1 i 0 1 \u00bc RtiM R1 i RiM : \u00f013:26\u00de Also, the following equality takes place m1 i \u00bc \u00f0m1 V \u00de2: \u00f013:27\u00de 13.3.4 Power of a Power-Load Element Let us consider the circuit with a variable power of power-load element P in Fig. 13.12. For its different values, we get the load power characteristic as a bunch of hyperbolas. The voltage source characteristic is the known straight line. 372 13 Power-Source and Power-Load Elements For example, this line intersects the hyperbola P1 into points M1, ~M1. Also, the characteristic point M \u00fe corresponds to the tangent hyperbola PM . From (13.23), we get PM \u00bc \u00f0V0\u00de2 4Ri ; VM \u00bc V0 2 : \u00f013:28\u00de We must prove the single-valued area of the power-load element characteristic. To do this, we may use expression (13.24). In this case, we get the similar form P \u00bc V0V Ri \u00f0V\u00de2 Ri : \u00f013:29\u00de This dependence P\u00f0V\u00de determines a parabola in Fig. 13.13. Therefore, the single-valued mapping of the parabola points onto the axis V takes place, which coincides with Fig. 13.11. So, we may use the same cross-ratio (13.25) for the initial point V1 at once m1 V \u00bc VM V1 V0 1 \u00bc V1 V0=2 V0=2 : \u00f013:30\u00de The conformity of the points V1, m1 V is shown in Fig. 13.13. P M P 1 V V M M + M 1 V 0 R i VV1 M 1 I P + V 0 0 Fig. 13.12 Characteristics of a power-load element with different values of its power 13.3 Influence of Voltage Source Parameters and Power-Load Element \u2026 373 Similarly to the above, let us consider the cross-ratio m1 P for the power P1 using the conformity of the variables by Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000556_20140824-6-za-1003.01018-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000556_20140824-6-za-1003.01018-Figure3-1.png",
        "caption": "Fig. 3. The grasp G producing the wrench set W = {W1,1,W2,1,W3,1} is non FC. H1 andH2 are the hyperplanes that fail the force closure test, and then W2,1 indicates the point to be replaced. The replacement point lie in region bounded by the separating hyperplanes H \u2032 1 and H \u2032 2, passing through the origin and parallel to H1 and H2, respectively. Choosing the contact point associate to the wrench W \u2217 in this region, the new obtained grasp is FC.",
        "texts": [
            " It starts generating a random grasp G; if G is not FC then a iterative replacement of contact point in G is executed. The contact points to be replaced are those whose wrenches define boundary hyperplanes H of CH(W ) that do not satisfy the condition of the forceclosure test describe in the previous section. The points for the replacement are those whose wrenches belong to the region of the wrench space bounded by separating hyperplanes H \u2032 containing the origin O and parallel to the hyperplanes H . Figure 3 illustrates a replacement for an hypothetical 2D wrench space. Step (3) simply forms the corresponding wrench set W l for the current considered grasp Gl. In Step (4) the algorithm searches the facet of CH(W l) closest to the origin, hereinafter FQ, and calculates the distance from FQ to the origin, i.e. the quality of the current grasp Ql. In Step (5) a subset \u2126l C with candidate points to replace one of the points that define FQ is built. This subset is determined employing separating hyperplanes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001499_memsys.2014.6765680-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001499_memsys.2014.6765680-Figure5-1.png",
        "caption": "Figure 5: Fabrication schematics of hierarchically wavy pattern",
        "texts": [
            " Figure 4 shows the micrograph of the reflowed microstructures on PDMS (SEM, Hitachi S-3000H). The reflowed structures were to ensure a continuous strain profile on the PDMS surface upon pre-straining so that the wrinkling structures were expected to be continuous. The PDMS replica was stretched uniaxially with the strain (40%) applied at 90\u00ba to the longitudinal direction of the reflowed microstructures. It was then subjected to the electrical discharge assisted surface wrinkling process with 30s exposure time (Figure 5). The resulting hierarchically wavy pattern is shown in Figure 6. Nanowrinkling pattern formed both in the grooves and on the ridges of the reflowed microstructures with fairly good regularity. The nanowrinkles oriented parallel to the longitudinal axis of the reflowed microstructures, demonstrating an evident topographical hierarchy. Red profile traces revealed that the nanowrinkles propagated along the curvilinear shape of the ridge of the microstructure, which contributed to a hierarchically wavy surface pattern"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000442_055013-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000442_055013-Figure2-1.png",
        "caption": "Figure 2. The toy in instantaneous profile: the sense of 3\u03c9 depends on the sense of \u03a9, the angular velocity of the symmetry axis about the direction of gravity.",
        "texts": [
            " Our analysis is similar to the treatment of coin whirling in [1]. In particular, we consider motion in which the centre of mass remains at rest while the toy rolls and spins without slipping. We take each ball to have mass M and radius R. We choose the origin to be the stationary centre of mass at the point of contact of the two balls. The torque is then due to the normal force of the surface. Assuming that the toy is supported by the surface, the normal force is Mg2 . The equation of motion of the toy is L t d d . (1)\u03c4 = Referring to figure 2 to define the angle \u03b8 used to describe the inclination of the toy, we see that the torque magnitude is Mg OA MgR2 2 sin\u03c4 \u03b8= = . We can also use figure 2 to understand the consequence of the non-slip rolling constraint. Two angular velocities will be crucial in understanding the motion. The angular velocity of the toy about its symmetry axis is 3\u03c9 . The symmetry axis itself rotates about an axis parallel to gravity and passing through the centre of mass of the toy. The angular velocity of the symmetry axis about the direction of gravity is called \u03a9. The total angular velocity of the toy is . (2)3\u03c9 \u03a9 \u03c9= + The lower ball traces a circle on the ground centred on the point where the vertical axis \u03a9 passing through the centre of mass intersects the ground at O. The diameter of this circle is AB R2 sin \u03b8= . The locus of points of the lower ball which successively contact the ground to trace this circle themselves form a circle on the ball, shown in profile (figure 2) with antipodal points A and A\u2032. This circle is the intersection of a conical surface with its apex at the ball\u2019s centre and aperture 2\u03b8. The diameter of this circle is A A R2 sin \u03b8\u2032 = . Since A A AB\u2032 = , the circumerences of these circles are equal and the no-slip condition implies 3\u03a9 \u03c9= . This constraint is the essential difference between a standard top and hurricane balls. The validity of the no-slip assumption is tested in the video available as an online supplement to this paper. A stroboscope is used to freeze the motion of a whirling hurricane balls toy, thus providing a measure of \u03a9",
            " In fact the tape moves, but only very slowing in comparison with the rotational frequency of the toy (about 40\u201350 Hz). We find that \u03a9 differs from 3\u03c9 by only a hertz or less. At least part of the slight failure of the no-slip assumption is likely due to the necessity of corralling the hurricane balls on a spherical surface to prevent translational drifting. Two surfaces of different curvature are tried. The smaller radius of curvature results in a greater difference between \u03a9 and 3\u03c9 . This can be understand by visualizing figure 2 with a spherical supporting surface. Under that geometry the aperture of the cone discussed above does not quite equal 2\u03b8. Air resistance and the Magnus force might also contribute to the slight slipping. Regardless of such complications, however, the no-slip assumption is seen to be a good approximation. In addition to the symmetry axis, we adopt body fixed axes 1 and 2 such that 1 2 3\u2212 \u2212 define a right-handed coordinate system at the centre of mass. In figure 2, axis 2 is instantaneously directed toward the reader while 1 lies in the page perpendicular to 2 and 3. These are principal axes. The angular momentum of the toy is L I \u00b7 \u03c9= , where I is the rotational inertia tensor. In our body fixed system, I is represented by MRI 14 5 0 0 0 14 5 0 0 0 4 5 . (3)2 \u239b \u239d \u239c\u239c\u239c\u239c\u239c\u239c \u239e \u23a0 \u239f\u239f\u239f\u239f\u239f\u239f = It is useful to express \u03a9 as the sum of projections along the principal axes. Using the geometry of figure 3 we write , (4)3\u03a9 \u03a9 \u03a9= +\u22a5 where sin\u03a9 \u03a9 \u03b8=\u22a5 and cos3\u03a9 \u03a9 \u03b8= . The total angular momentum then has the three contributions I\u00b7 I\u00b7( ) ( )3 3 3\u03a9 \u03c9 \u03a9 \u03a9 \u03c9+ = + +\u22a5 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002454_1350650116652566-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002454_1350650116652566-Figure3-1.png",
        "caption": "Figure 3. Unbalanced load calculation model.",
        "texts": [
            "a R2 p into equation (2), the dimensionless form of equation can be expressed as follows @ @ H 3 @ p @ \u00fe R L 2 @ @l H 3 @ p @l \u00bc @H @ \u00fe 2 @H @t \u00f03\u00de where p is the dimensionless oil film pressure, H is the dimensionless oil film thickness, l is the dimensionless axial coordinate, t is the computed time, is the fluid dynamic viscosity, L is the bearing width, and !a is the axis rotating speed. Motion equations of axis under dynamic loading When sliding bearing rotates, it will be subjected to dynamic loading. Dynamic loading are of many kinds, such as periodically unbalanced load and transient load generated from rubbing, shock, and vibration force. Although situations are different, the axis center O2 is not fixed and varies with time under dynamic loading. In Figure 3, O is the bearing center, O2 is the axis center that is denoted by (x, y), axis rotates counterclockwise with !a, Wx and Wy are the oil film forces for vertical direction x and horizontal direction y separately, and Mg is the weight of the axis. Force balance equations for horizontal direction and vertical direction are as follows M \u20acx \u00bc Wx\u00f0!at\u00de \u00feQx \u00feMg M \u20acy \u00bcWy\u00f0!at\u00de \u00feQy \u00f04\u00de Taking into account the uneven load of the rotating parts, the status of load is as shown in Figure 3. Od is the center of axis mass, ed is the eccentricity where the center of axis mass is relative to axis. The unbalanced amount Qx and Qy generated by the centrifugal force are as follows Qx \u00bcMed! 2 a cos!at, Qy \u00bcMed! 2 a sin!at \u00f05\u00de Introducing equation (5) into equation (4), the dimensionless acceleration equation \u20acX, \u20acY can be expressed as \u20acX \u00bc 6 L M!a R c 3 Wx\u00f0!at\u00de \"d cos t g c!2 a \u20acY \u00bc 6 L M!a R c 3 Wy\u00f0!at\u00de \u00fe \"d sin t \u00f06\u00de where Wx and Wy are dimensionless oil film forces for the vertical direction x and horizontal direction y separately"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001782_iros.2015.7354084-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001782_iros.2015.7354084-Figure14-1.png",
        "caption": "Figure 14. Prototype walking support robot.",
        "texts": [
            " When the magnitude of the Shaft A\u2019s torque does not exceed the detection torque level, the torque is transmitted to Shaft B. If the magnitude of the torque exceeds the detection torque level, then the torque limiter cuts off the torque transmission, Plate Z sides along Shaft A and switches off. The detection torque level is adjustable by using the torque limiter\u2019s adjusting nut. We developed the prototype walking support robot equipped with velocity, torque, and contact force-based mechanical safety devices. Fig. 14 shows the developed robot. As shown in Fig. 14, the length is 146 cm and the width is 154 cm. The armrest is adjustable in height from 85 cm to 108 cm according to the height of each elderly person by using a hand crank. Contact Force Detection Plate Compression Spring Slide Rail Switch E Velocity-based Safety Device A Forward Figure 11. Contact force-based safety device. (c) (a) (b) (d) Switch E Contact Force Detection Plate F Linear Spring E Pin E Pawl E Wire B Pully Linear Spring F Plate A Figure 12. Mechanism of the contact force-based safety device"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003238_detc2016-59654-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003238_detc2016-59654-Figure4-1.png",
        "caption": "Figure 4 Tooth tip relief",
        "texts": [
            " This leads to a computationally efficient gear dynamics simulation suited for the design optimization procedure considering wind load uncertainty. For gear design optimization to ensure expected service life under wind load uncertainty, an integrated numerical procedure is developed using the wind uncertainty model, the pitting fatigue prediction model and multibody gear dynamics simulation procedure discussed in the previous sections. For given design variables d including gear tooth geometry parameters such as face width and tip relief (Fig.4), the random time-domain wind speed data is generated for a given 10- minute mean wind speed V10 and turbulence intensity I10 characterized by the averaged joint PDF using NREL TurbSim [31]. The wind speed data is then inputted into NREL FAST [32] to perform the time-domain coupled nonlinear aero-hydroservo-elastic simulation of a wind turbine considering the pitch control of the rotor blade. The transmitted torque and speed of the rotor hub are predicted and used as input to the high-fidelity multibody gear dynamic simulation for calculating the 10- minute mesh force variation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000028_s0219455411004038-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000028_s0219455411004038-Figure2-1.png",
        "caption": "Fig. 2. The \\real\" con\u00afgurations of the membrane element.",
        "texts": [
            " The orientation of the principal directions is calculated using the elementary second-order tensor transformation. This transformation uses the local pure stress components of the element, x; y, and xy. The stresses are obtained using the following linear constitutive relation: \u00be \u00bc D\"; \u00f08\u00de where \u00be comprises the stress components x, y and xy, and \" comprises the pure strain components \"x, \"y, and xy. The pure strains are the elastic strains only, i.e., free of rigid body components. These pure strains are uniquely obtained using the following geometry-based relations associated with Fig. 2. \"x \u00bc AC 0 AC AC \u00f09\u00de \"y \u00bc B 0D 0 BD BD \u00f010\u00de xy \u00bc \\DBD 0 \\DBD 00 \u00bc AD 0 AD BD \u00f0AC 0 AC\u00deAD AC BD ; \u00f011\u00de where xy is de\u00afned as the angle change from a right angle which would have been if C would have stayed in place and not moved to C 0. The e\u00aeect of C moving to C 0 is like starting from D 00 instead of D. Thus, xy is the pure shear strain. With the local stress components given in Eq. (8), the angle of the principal plane with respect to the local coordinate system is calculated as: tan\u00f0 local\u00de \u00bc 2 xy x y 2 ; \u00f012\u00de where local is the angle between the local x-direction and the principal direction 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.5-1.png",
        "caption": "Fig. 3.5 Positive definition of a internal reactions and b rotation (but negative slope)",
        "texts": [
            " The vertical position of a point with respect to the center line of the beam without action of an external load is described through the z-coordinate. The vertical displacement of a point on the center line of the beam, meaning for a point with z = 0, under action of the external load is indicated with uz. The deformed center line is represented by the sum of these points with z = 0 and is referred to as the bending line uz(x). In the case of a deformation in the x\u2013z plane, it is important to precisely distinguish between the positive orientation of the internal reactions and the positive rotational angle, see Fig. 3.5. The internal reactions at a right-hand boundary are directed in the positive directions of the coordinate axes. Thus, a positive moment at a righthand boundary is clockwise oriented, see Fig. 3.5. However, a positive rotational angle is counterclockwise oriented, see Fig. 3.5. This difference requires some careful derivations of the corresponding equations. Only the center line of the deformed beam is considered in the following. Through the relation for an arbitrary point (x, uz) on a circle with radius R around the center point (x0, z0), meaning (x \u2212 x0) 2 + (uz(x) \u2212 z0) 2 = R2, (3.1) one obtains through differentiation with respect to the x-coordinate 2(x \u2212 x0) + 2(uz(x) \u2212 z0) duz(x) dx = 0, (3.2) alternatively after another differentiation: 2 + 2 duz dx duz dx + 2(uz(x) \u2212 z0) d2uz dx2 = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001797_humanoids.2015.7363511-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001797_humanoids.2015.7363511-Figure1-1.png",
        "caption": "Fig. 1. Balance recovery strategies: (a) Postural balance strategy and (b) reactive stepping strategy.",
        "texts": [
            " Abdallah and Goswami [2] developed a postural balance controller that controls the rate of change of linear and angular momenta of a humanoid robot. More recently, Micchietto et al. [4] proposed a whole body postural balance controller that identifies the desired CoP as the high level input. Lee and Goswami [7] presented a balance controller for a non-level and non-stationary ground. Among a number of strategies that a humanoid robot can employ to maintain balance, two most representative strategies are the postural balance strategy and the reactive stepping strategy as shown in Fig. 1. The postural balance strategy is usually chosen for relatively short and mild perturbations, against which a humanoid robot can maintain balance in place simply by controlling the ankle or hip joints [8], [9], or by rotating the whole upper body [2]\u2013 [5], [7]. Upon the long or strong perturbations, a humanoid robot should take a reactive stepping strategy in which a robot takes one or more steps to prevent falling [10]\u2013[13]. Recently, researchers have applied momentum control approaches to reactive stepping"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000495_icit.2015.7125212-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000495_icit.2015.7125212-Figure1-1.png",
        "caption": "Figure 1- a) Microbial fuel cell with acetate as a substrate. With the substrate oxidation, electrons flow from the anode to the cathode through the external circuit. Protons rising from the oxidation reaction are then selectively chosen to go to the cathode chamber through a proton exchange membrane; b) CAD representation of the reactor geometry; c) the actual assembled reactor.",
        "texts": [
            " Cation exchange membranes, anion exchange membranes and bipolar membranes have commercial availability. Of all three, the cation exchange membrane has the lowest resistance and is a usual choice [22], [20], [26], [28]. As for the bacteria and substrate, wastewater is usually a good source for both and needs no additions. The study and characterization of a MFC\u2019s electrical behaviour is usually conducted with the use of a variable resistance, a voltmeter and an amperemeter. The MFC constructed by the authors is pictured at the most right on Figure 1 with a typical schematic at the most left and the CAD model in the middle. Each reactor has an approximate empty bed volume of 1.5 dm 3 . The power production capability of the cell was evaluated using acetate as the medium for microbial anaerobic organisms grown at the anode, a water aerated chamber for the cathode and a Nafion PEM membrane. A period of roughly 2 days was provided for reactor acclimation. A set of 14 voltage values was retrieved and used for building the polarization curve. Experimental data is plotted on Figure 2: high external resistance values limit current output and voltage production between the module\u2019s terminals (Figure 2a))",
            " However, they are expensive and not the best solution for studying the cell\u2019s electrical behaviour since their complexity and operating method are more adequate for biochemical studies. Manually varying a resistance load, applied to the MFC, and measuring its voltage and current, is another solution. Figure 4 \u2013 Application programming interface developed with Processing. The interface is composed of a configuration console and a display window. Figure 3 \u2013 Experimental measurement and test platform. Such a method was previously used to characterize the MFC shown in Figure 1 to achieve the results presented in Figure 2. The use of the described method turned to be cumbersome and time consuming. The use of an automated platform, capable of digitally varying the load values while simultaneously measuring biochemical parameters can answer to the shortcomings of both the identified solutions. The platform was developed around an ARM microprocessor, namely the mbed LPC1768 (see Figure 3). A PC application sends configuration data to the mbed which, in turn, decodes the data to configure the biochemical sensors and the load varying assembly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000212_978-981-15-5712-5-Figure25-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000212_978-981-15-5712-5-Figure25-1.png",
        "caption": "Fig. 25 The Industrial Revolutions indicated by the changes in slope of the GDP growth rate",
        "texts": [
            " It has been mentioned earlier that the invention of prime movers led to an Industrial Revolution\u2014called the First Industrial Revolution. If production of iron is taken as a measure of the degree of industrialization, then Fig. 24 shows the characteristic change in the growth rate of Iron production. The sudden change in the slope indicates a paradigm change and the Industrial Revolution. The Second Industrial Revolution (IR) is considered to be caused by the emergence of computer technology and the semiconductor industry. This can be identified with the help of the economic growth characteristics shown in Fig. 25. The primary impetus to the Second Industrial Revolution came from the miniaturization of electronic devices. Figure 25 shows mainly two points where the slope changes and, so, the author feels that there have been really two revolutionary changes in the industry. Some historians have made finer 22 A. Ghosh divisions and suggested more numbers of IR. The trend of miniaturization in engineering activities can be observed in the trend of development shown in Fig. 26. There are a number of reasons behind this trend of miniaturization. Ability to incorporate much higher level of intelligence in the machines and devices is one of the reasons",
            " The possibilities for design are almost unlimited with electrothermal actuation. It paves the way for topology optimization involving multiphysics simulation [31\u2013 33]. Analysis of electrothermal actuators is challenging because three simulations\u2014 electrical, thermal, and elastic\u2014are to be performed in a sequence. Conduction, 48 G. K. Ananthasuresh convection, and radiation effects as well as temperature-dependent properties are to be accounted for in simulation and design [34]. Both intuition and systematic optimizations can lead to interesting designs. Shown in Fig. 25 is a planar robotic platform that uses three arrays of expanding parallel heatuators. The bent-beam actuator and heatuator building blocks are used in arrays or in special arrangements. We see them in many actuators that use electrothermal principle. They are thus ubiquitous\u2014the same principle used in many places. The Art and Signs of a Few Good Mechanical Designs in MEMS 49 9 Modularity in a Compliant xy-Stage Many designs illustrated so far have modular construction: a building block repeated in different ways in a specific arrangement to serve a function that is beyond what the building block does"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002652_978-3-319-33714-2_16-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002652_978-3-319-33714-2_16-Figure2-1.png",
        "caption": "Fig. 2 Schematic model of the underactuated manipulator with passive joint (left) and reference trajectory of the end-effector r d (right)",
        "texts": [
            " In addition, the method provides a way to quantify the effect of each fuzzy-valued parameter on the uncertainty of the system output by influence measures, see Gauger et al. (2008). Providing systematical tools to analyze the influences of parameter uncertainties, the fuzzy-arithmetical analysis allows a better understanding of the dynamical behaviour of mechanical systems, successfully applied to a flexible parallel manipulator, see Walz et al. (2015). The proposed adaptive nonlinear control scheme is applied to an underactuated manipulator with passive joint, depicted in Fig. 2. The manipulator consists of three rigid beams with length l1 = 1m and l2,3 = 0.5m, of which the first and second beam are driven by motor torques T1 and T2, whereas the third beam is linked to the second beam by a torsional spring-damper combination. Introducing the generalized coor- dinates qT = [ \ud835\udefc \ud835\udefd \ud835\udefe ] and the input uT = [ T1 T2 ] the equation of motion (1) can be derived. In order to investigate the impact of uncertainties on the control structure, the system parameters are described as symmetric triangular fuzzy numbers with the membership function \ud835\udf07(x) = \u23a7\u23aa\u23a8\u23aa\u23a9 1 + (x \u2212 p\u0302i)\u2215(\ud835\udefcip\u0302i), for (1 \u2212 \ud835\udefci)p\u0302i < x \u2264 p\u0302i, 1 + (x \u2212 p\u0302i)\u2215(\ud835\udefcip\u0302i), for p\u0302i < x < (1 + \ud835\udefci)p\u0302i, 0, otherwise, (14) where p\u0302i denotes the nominal value of the ith parameter and \ud835\udefci describes its deviation and are given in Table 1. To track a desired trajectory r d , see Fig. 2, the end- effector position r is approximated by the linearly combined output yT = [ \ud835\udefc \ud835\udefd ] +[ 0 l3\u2215(l2 + l3) ] \ud835\udefe by solving the inverse kinematics. To demonstrate the need of an adaptive control scheme, a standard control approach consisting of feedback linearization and asymptotic tracking control v = y\u0308 d + k1(y\u0307d \u2212 y\u0307) + k2(yd \u2212 y) is compared with the proposed control scheme consisting of both non-adaptive and adaptive MPC. The fuzzy-valued simulation results of the tracking errors e y , depicted in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002416_s00033-016-0659-6-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002416_s00033-016-0659-6-Figure2-1.png",
        "caption": "Fig. 2. Inverted triple pendulum under harmonic motion of the support",
        "texts": [
            " Since the equations are coupled, the static and dynamic components interact in determining stability. By substituting Cj = R ei\u03c3t, where R = u+iv, we can recast the system (35) in matrix form, similarly to the resonant case \u03a9 = \u03c9/2 (for which coupling also occurs). The matrix H now will be: H = \u239b \u239c \u239c \u239d 0 1 0 0 \u03b12 \u03b11 2\u03b13 2\u03b14 \u03b16 \u03b15 \u03b19 \u03b110 \u2212 \u03b18 + \u03c3 \u03b17 0 \u03b110 + \u03b18 \u2212 \u03c3 \u03b19 \u239e \u239f \u239f \u23a0 . (36) The stability conditions have the form of inequalities (17). As a sample system we will consider a triple pendulum in its unstable upright equilibrium position (Fig. 2). The system is made of three hinged rigid rods of equal length l, elastically restrained at the hinges by linear springs of equal stiffness c, carrying heavy masses m at the hinges. Hinges are damped by linear viscous devices of equal constants d. The support undergoes a vertical harmonic motion z = a cos \u03a9t, with a l. By taking the rotations qi (i = 1, 2, 3) as Lagrangian coordinates, the equations of motion were derived by using Lagrange equations, with F = 1 2d [ q\u03072 1 + (q\u03072 \u2212 q\u03071) 2 + (q\u03073 \u2212 q\u03073) 2 ] the dissipation function"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure7.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure7.3-1.png",
        "caption": "Fig. 7.3 Regime change at the expense of conductivity yN ! yN",
        "texts": [
            "6) I21 \u00bc 1:6 0:974 0:04 0:974\u00fe 0:0648 0:82\u00fe 0:09396 1:61\u00fe 1 \u00bc 1:5592 1:2433 \u00bc 1:254; I22 \u00bc 2:0538 0:82 1:2433 \u00bc 1:356; I23 \u00bc 2:3394 1:61 1:2433 \u00bc 3:1: 212 7 Recalculation of Load Currents of Active Multi-ports 7.2 Recalculation of Currents for the Case of Changes of Circuit Parameters We consider the concrete circuit (see Sect. 6.1.3) in Fig. 7.2. The above results allow investigating the influence of a lateral conductivity yN and longitudinal conductivity y0N on load currents [2] 7.2.1 Change of Lateral Conductivity Let the conductivity value yN be changed, yN ! yN . Corresponding changes of an initial point M ! M and short circuit point SC ! SC are shown in Fig. 7.3. On the other hand, the coordinates of the points G1; G2 by (6.26) and (6.24) do not depend on the value yN . Therefore, a straight line G1 G2 is the fixed line. The point 0, as the open circuit regime, does not depend on this element too. According to Fig. 7.1, the influence of this conductivity yN can be interpreted as a projective transformation in the plane I1; I2. This transformation, as similar to 7.2 Recalculation of Currents for the Case of Changes of Circuit Parameters 213 (7.2) and (7",
            " Therefore, the cross-ratio has the same value for the points ISC1 ; ISC1 mN \u00bc \u00f0 0 I1 I1 IG11 \u00de \u00bc \u00f0 0 ISC1 ISC1 IG11 \u00de \u00bc I1 0 I1 IG11 I1 0 I1 IG11 \u00bc ISC1 0 ISC1 IG11 ISC1 0 ISC1 IG11 : \u00f07:12\u00de Now, let us formulate the value mN by yN ; yN and determine the sense of the parameter mN . Viewing expressions (6.28), we may consider expression (7.8) of mN as a non-uniform coordinate mN \u00bc ISC1 ISC1 dSC3 dSC3 \u00bc ISC1 dSC3 ISC1 dSC3 \u00bc n13 n13: Hence, the value mN is the cross-ratio of the initial SC and subsequent points SC shown in Fig. 7.3. The base values of conductivity yN are values y0N ; y G1 N . The value y0N \u00bc 1 determines the current I1 \u00bc 0. The value yG1N presets the current I1 \u00bc IG11 as I2 \u00bc 0. Analysis of the circuit gives the following relationship for yG1N yG1N \u00bc y0N \u00fe y1y1N y1 \u00fe y1N \u00fe y2y2N y2 \u00fe y2N \u00bc yiN : \u00f07:13\u00de The sense of value yiN will be explained below. Therefore, the cross-ratio has the view mN \u00bc \u00f01 yN yN yiN \u00de \u00bc yN 1 yN \u00fe yiN yN 1 yN \u00fe yiN \u00bc yN \u00fe yiN yN \u00fe yiN : \u00f07:14\u00de Now, we formulate the value mN by the yN ; yN for the general case of currents I1; I2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002236_icsec.2014.6978198-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002236_icsec.2014.6978198-Figure1-1.png",
        "caption": "Fig. 1. Motion sensor attachment and data collection.",
        "texts": [
            " According to the previous related works of golf swing pattern analysis, this paper proposes the idea of the data classification for particular golf swing patterns and the suitable analysis of swing characteristics which will be mentioned in the next Section. III. THE DESIGN OF GOLF SWING PATTERN ANALYSIS AND ALGORITHMS This research used motion sensors of 9 Degrees of Freedom - Razor IMUs to detect body movements [12]. The sensors were attached to the upper and lower back to observe the muscle movement of the upper back, lower back, and spinal axis as shown in Fig. 1. As soon as the golf player made a golf swing, the swing data were transferred at a real time through Bluetooth to the connected computer. The collected data were processed by using the classification program to classify and compare the golf swing data. The detail will be described in topic C and D. Two types of golf swing data: Acceleration and Rotation were transferred from Razor IMUs. Acceleration data were measured in the range of \u00b1 1 6 g in 3D plane (X, Y, and Z). Rotation data were measured in the range of \u00b1180 in 3D plane (Roll, Pitch, and Yaw) and the data were sent at 50 times per second as shown in Fig. 1. The data transformation was processed in 2 steps: deviation reduction and data classification using the similar changing ratio at each period of time. The deviation reduction covered the un-stability of muscle spasticity and muscle relaxation. Consequently, the measurement of swing data appeared having uncertainty variance. Therefore, the average data at a certain time was calculated, as in (1) where k is the time sequence n is range of time to average OD is real data collected from sensors ND is adjusted data After reducing the data deviation, new data (ND) were classified by the algorithm in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.36-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.36-1.png",
        "caption": "Fig. 2.36 Plane truss structure. Nodes are symbolized by circles (\u00a9)",
        "texts": [
            " State for this problem three boundary conditions and the appropriate differential equation under the assumption that the Young\u2019s modulus E is a function of the spatial coordinate x . \u2022 Which general types of \u2018load conditions\u2019 did we distinguish for rod problems? \u2022 State the required (a) geometrical parameters and (b) material parameters to define a rod element. \u2022 Sketch the interpolation functions N1(x) and N2(x) of a linear rod element. \u2022 State four (4) characteristics of a finite element stiffness matrix. \u2022 The stiffness matrix for a rod element can be stated as (Fig. 2.36) K e = E A L [ 1 \u22121 \u22121 1 ] . 2.5 Supplementary Problems 75 Fig. 2.35 Axially loaded continuum rod Which assumptions does this equation involve in regards to the (a) material and (b) the geometry? \u2022 State the DOF per node for a truss element in a plane (2D) problem. \u2022 State the DOF per node for a truss element in a 3D problem. \u2022 The following Fig. 2.36a shows a plane truss structure which is composed of 15 rod (E, A) elements. State the size of the stiffness matrix of the non-reduced system of equations, i.e. without consideration of the boundary conditions. What is the size of the stiffness matrix of the reduced system of equations, i.e. under consideration of the boundary conditions? Consider now Fig. 2.36b where the rod element 1\u20132 (length L) has been replaced by a spring of stiffness k = E A L . How does the overall stiffness of the truss structure change? 2.9 Simplified model of a tower under dead weight (analytical approach) Given is a simplified model of a tower which is deforming under the influence of its dead weight, cf. Fig. 2.37. The tower is of original length L , cross-sectional area A, Youngs modulus E , and mass density . The standard gravity is given by g. Calculate \u2022 The reduction in length L = L \u2212 ux (L) due to acting dead weight"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001025_s12647-014-0101-5-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001025_s12647-014-0101-5-Figure2-1.png",
        "caption": "Fig. 2 Transformation of coordinate systems for deducing the output acceleration of a vibrafuge",
        "texts": [
            "com rotation matrices are required to determinate the displacement of a rigid body, and we get the acceleration by the second-order derivative of the displacement. A transformation matrix is in the form of M \u00bc A D 0 1 \u00f01\u00de where A is a 3 9 3 normal matrix, which represents the attitude change of a coordinate system (CS) caused by clockwise rotation and D is a 3 9 1 translation vector, which describes the relative position of the CS to its initial state. Next, we created ordinal CSs as indicated in Fig. 2. A reference CS is built as ox0y0z0. Axis oz0 is along the local gravity acceleration and directed upward. Axis oy0 is the initial position of the centrifuge arm and axis ox0 is confirmed by the right-hand rule. Frame ox1y1z1 is linked to the centrifuge, and frame ox2y2z2 is linked to the accelerometer under test. Other coordinate systems are auxiliaries. z0(z1) means oz0 and oz1 axes are in the same direction. Meanings of x21(x22), z2(z22), y1(y20, y21) can be inferred by analogy. Based on the CSs, the transformation process is as follows: (1) ox0y0z0 rotates around axis oz0 with angular velocity X to form ox1y1z1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001604_aieepas.1958.4499867-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001604_aieepas.1958.4499867-Figure8-1.png",
        "caption": "Fig. 8. Single-suspension bundled asembly with circular yoke-plate shield in various positions",
        "texts": [
            " 6 includes the addition of a circular 24-inch-diameter shield made from 11/4- inch iron-pipe-size aluminum tubing, used to compensate for the corona and RIV effect of the yoke plate. The re- cADcA, (A, (B) ,.i n0 0 x00 0 4 U, I1-i0 0 Ix (C) (D) Fig. 6. Single-suspension bundled-conductor assemblies with circular shield around yoke plate and clamp shields sults are illustrated by the curves in Fig. 7. Some consideration was given to the most effective location for the circularyoke-plate shield. Fig. 8 shows this shield arranged in several different positions and Fig. 9 shows an RIV comparison of these position changes. To fulfill the requirements of shielding the clamp, yoke plate, and other connecting hardware, as well as providing grading for bottom units of insulator strings, a corona shield, shown in Fig. 10, was designed and tested. The RIV curves, Figs. 11 and 12, demonstrate the effectiveness of this type of shield. Shielding of Double-SuspensionString Bundled Arrangements Frequently, transmission-line designs require double strings of insulators in suspension at specific locations",
            "05-inch-OD conductor, the ratio of conductor diameters in single versus bundled assemblies will exert greater influence on RIV characteristics than will the ratio of currentcarrying capacities. The RIV characteristics of two bundledconductor arrangements are shown in Fig. 3. Incieasing the conductor size from 1Kaminski, Jr.-Corona Shields for Suspension Assemblies 91APRIcL 1958 +-Yoke-plate shield, position (1), Fig. O-Yoke-plate shield, position (2), Fig. *-Yoke-plte shield, position (3), Fig. E Yoke-plate shield, position (4), Fig. 8 8 8 8 1.05 to 1.90 inches reduced the RIV at critical voltage 39%. At voltages higher than 250 kv, the RIV characteristics of the two arrangements are essentially the same. At these higher voltages the increased corona activity on the supporting hardware apparently overcomes the compensating effect of the increased diameter on RIV. The minimum corona points D shown in Figs. 2 and 3 are primarily due to corona activity off the U-bolts of suspension clamps. It follows, therefore, with proper shielding of suspension clamp, improvement in RIV performance can be realized",
            " The addition of clamp shields as shown in Figs. 6(B), (C), and (D) reduces the RIV level 82%. It is interesting to note that the appearance of corona has been entirely removed from these latter three assemblies throughout the range of test voltages except at the open ends of the circular yoke-plate corona shield at 277 kv. Changing the position of the yoke-plate shield with respect to the yoke itself has little effect on the RIV characteristics, particularly at critical voltage. This is demonstrated by varying the positions as shown in Fig. 8 with tbe associated curves of Fig. 9. A corona shield designed specifically for bundled-conductor applications is shown in Fig. 10 and with characteristic curves in Fig. 11, and the 1.9-inch-OD conductor assembly shows a reduction in RIV when compared with an unshielded assembly. This is equivalent to the reduction affected by yoke and dlamp shields when used simultaneously as illustrated in Fig. 6. No corona was observed on this shielded assembly at the maxm applied test voltage. An 84% reduction in RIV was obtained (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001520_801519-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001520_801519-Figure13-1.png",
        "caption": "Figure 13: Experiment topology.",
        "texts": [
            " When our local remapping algorithm is applied the increment ratio is decreased, but still higher than isotropic. As a consequence, when sensor nodes are deployed in three-dimensional space, it is better to deploy sensor nodes as flat as possible. If network height is much longer compared to surface side length, sensor nodes should be rotated 90 degrees to reduce power consumption caused by dipole antenna radiation pattern. We tested our algorithm in the real network. Because of limited node number, our experiment is composed of very simple topology as shown in Figure 13. There are three nodes in the network. Two nodes (node 2 and 3) are located in the ground (\ud835\udc65-\ud835\udc66 plane) and last node is located at the upper side of node 2. The distance between nodes 1 and 3 is same as the distance between nodes 2 and 3. Also nodes 1, 2, and 3 are located on the \ud835\udc65-\ud835\udc67 plane. We used micaz based modified sensor node. The sensor nodes have PCB antenna which shows almost the same radiation pattern with dipole antenna. The sensor node uses CC2420 RF chipset and the chipset provides variable output power from \u221225 dBm to 0 dBm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001518_2015-01-9087-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001518_2015-01-9087-Figure6-1.png",
        "caption": "Figure 6. Fluid film pressure distribution of a spiral grooved journal bearing",
        "texts": [
            " Radial load is a function of the pump geometry, discharge flow and fluid properties, and can be determine empirically from [18]: Misalignment in the bearing can be defined as being made up of two components as follows: (4) where, \u03b3 is the total misalignment of one bearing, \u03b3mfc is the misalignment due to the manufacturing process and assembly of the pump, \u03b3load is the misalignment caused by load obtained by solving the following equation: (5) where, L1 and L2 are the distances in mm. Once the load and misalignment are known, three dimensional Thermo hydrodynamic performance analyses are performed to calculate the pressure and temperature distribution. On the fluid film, wherever the pressure falls below the vapor pressure, it is equated to zero (Figure 6). The load carrying capacity is then calculated by surface integral of the pressure distribution using Equations 6, 7, 8: (6) (7) (8) where, FR and FT are the radial and tangential loads, W is the resultant reaction of the fluid film in the bearing and \u03a6 is the bearing angle. This procedure is repeated for different eccentricity ratios. The equilibrium eccentricity ratio is then calculated, when the reaction load developed by the fluid film (Equation 8) equals the load requirement for the bearing [18]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002864_0954405416661003-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002864_0954405416661003-Figure4-1.png",
        "caption": "Figure 4. Profile modification: (a) design of contact path and (b) pre-designed function of transmission errors.",
        "texts": [
            " So, for the developing profile-crowned gear, the pre-designed function of transmission errors at CORNELL UNIV on September 12, 2016pib.sagepub.comDownloaded from is performed simply by substituting the roll ratio with a specific function in the process of imaginary generation of gear. The function is represented by the following equation u2 =m21u1 +Du2(u1) \u00f018\u00de In the research, a fourth-order polynomial function is chosen as the pre-designed function of transmission errors Du2(u1) represented as follows Du2(u1)= b0 + b1u1 + b2u 2 1 + b3u 3 1 + b4u 4 1 \u00f019\u00de where bi(i=0 4) are the unknown coefficients. Figure 4 shows the procedures to pre-design a func- tion of transmission errors for Pm 2 : (1) designing the contact path lt through the initial point M (in the projection plane) and providing the bias angle h of lt that is inclined with respect to local reference axis H (Figure 4(a)). (2) Calculating the distance lti from point i(i=A,B,M,C,D) to the initial point M computationally in the direction of contact path lt, respectively. (3) Assigning magnitude of transmission error Du2i at each point i that is illustrated in Figure 4(b). From equation (16), we know that the parameter u1 may be represented by the parameters u1 and u1 linearly. So, the parameter u1 can be omitted while the fully conjugated tooth surface Pm 2 is expressed as two parameters (u1, u1). Therefore, the position and unit normal vectors of surface Pm 2 may be denoted by rm2 (u1m, u1m) and nm2 (u1m, u1m), respectively. Following the same steps, the modified tooth surface Pn 2 of the profile can be yielded by using the specific function of roll ratio described in equation (18) in the process of imaginary generation of gear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001799_pola.28032-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001799_pola.28032-Figure5-1.png",
        "caption": "FIGURE 5 ssDNA surface coverage (left column) and HE2 (right column) on SPAN-GNO nanocomposites with mole ratios of ANI to ABSA (1:0, 3:1, 3:2, 1:1, 2:3, 1:3, and 0:1). Each point is the mean of three measurements, and the error bars correspond to the standard deviation. [Color figure can be viewed in the online issue, which is available at wileyonlinelibrary.com.]",
        "texts": [
            " At the same time, the DPV signals of MB recording for pDNA immobilization and hybridization at other nanocomposites modified electrodes display the same tendency [Fig. 4(B\u2013G) and Supporting Information S4 Fig. S10A-C, S11] The ssDNA surface coverage (left column; for details of calculation process, see S5\u2013S7) and hybridization efficiency (right column; HE25 (IssDNA \u2013 IdsDNA)/IssDNA 3 100, where IssDNA represents current before hybridization and IdsDNA represents current after hybridization) on various electrodes constructed with different mole ratios of SPAN-GNO nanocomposites were respectively investigated and revealed intuitively in Figure 5. It is obvious that ssDNA surface coverage and HE2 change along with the changing of mole ratio of ANI to ABSA. One has to be noted that ssDNA surface coverage and HE2 reach the maximum when mole ratio change to 2:3. Then, 2:3 can be chosen as the optimal mole ratio of ANI to ABSA. In addition, Supporting Information S4 Figure S10D and S12 intuitionally reflect the effect of composition and reaction time on ssDNA surface coverage and HE2 through histogram. Meanwhile, the relevant electrochemical data produced as measured by DPV technique (Supporting Information S4 Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.42-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.42-1.png",
        "caption": "FIGURE 8.42",
        "texts": [
            " These requirements are satisfied within individual flume sections but not necessarily across sections. The initial conditions, including initial position and velocity of the riding object, must be provided to solve the equations of motion: u 0 \u00bc u0;w 0 \u00bc w0; _u 0 \u00bc _u0; and _w 0 \u00bc _w0 (8.117) The system of ordinary differential equations can be solved numerically for positions u(t) and w(t), velocities _u\u00f0t\u00de and _w\u00f0t\u00de, and accelerations \u20acu\u00f0t\u00de and \u20acw\u00f0t\u00de of the riding object using, for example, Wolfram\u2019s Mathematica (1998). Awaterslide configuration consisting of 20 flume sections, shown in Figure 8.42, was modeled and analyzed (Chang 2008). The overall size of the waterslide was about 300 in. 1150 in. 378 in. Note that the riding object started at the center of the cross section (w \u00bc 0.5) of the top section. The friction coefficient was assumed to be m \u00bc 0.08. The path of the riding object can be seen in Figure 8.42, which shows the object running over the edge of the flume section at three critical areas A, B, and Cwhich posed a safety hazard to the rider. The design had to be revisited by either changing the composition of the configuration or using closed-flume (360 deg.) instead of open-flume sections (180 deg.) currently employed. The overall riding time was 21.2 seconds which was very close to what was reported by the company that designed the waterslide (20 seconds). The maximum acceleration and velocity were 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure15-1.png",
        "caption": "Fig. 15 Force components affecting bending moment in a femur, b tibia",
        "texts": [
            " Shear forces in the bone sections can be derived based on the transverse components of the forces acting on the bones, as presented in Fig. 14. Again, two sections can be distinguished in each bone with the center of mass as the common boundary. Shear forces T1 in the femur and T2 in the tibia are described by the following equations. T1 = { Fhy1 x1 \u2264 Cm1L1 Fhy1 + m1gy1 x1 > Cm1L1 (43) T2 = { Fky2 x2 \u2264 Cm2L2 Fky2 + m2gy2 x2 > Cm2L2 (44) Shear forces and moments applied to the bone are the contributors to the bending moment inside the bone as illustrated in Fig. 15. The bending moment equations can be formulated based on graphical representation in Fig. 15 as follows. M1 = { \u2212Mh + Fhy1x1 x1 \u2264 Cm1 L1 \u2212Mh + Fhy1x1 + m1gy1(x1 \u2212 Cm1 L1) x1 > Cm1 L1 (45) M2 = { \u2212Mk + Fky2x2 x2 \u2264 Cm2 L2 \u2212Mk + Fky2x2 + m2gy2(x2 \u2212 Cm2 L2) x2 > Cm2 L2 (46) In contrast to the normal and to the shear forces that are constant within sections, the bending moment is expressed as piecewise linear functions with a nonzero slope coefficient. It can be noticed that the forces and moments applied at the right end of the tibia and femur are not considered in the equations. This is because the bone has to be in equilibrium and the value of forces and moments computed at the right end from derived equations will be equal in magnitude to the forces applied at that end"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.11-1.png",
        "caption": "Fig. 3.11 Pin and spur surfaces of a five-point wheel suspension",
        "texts": [
            " Contact between two bodies can occur on two non-physical, spatially fixed surfaces or rather body fixed surface, spur surface or rather pin surface, along the instantaneous axis of rotation or rather screw axis, which indicate the instantaneous motion state. In the suspending motion of a five-link wheel suspension, for example, the motion of the wheel carrier relative to the chassis can be represented as a screw motion of the wheel carrier-fixed pin surface with respect to the chassis-fixed spur surface. The contact line is the instantaneous screw axis of the spatial motion of the wheel carrier with respect to the chassis (Fig. 3.11). An arbitrary, spatial kinematic chain, with or without considerations of kinematics consists of \u2022 nB bodies (without a reference body), \u2022 nG joints, each having \u2022 fGi joint DoF, where the individual body, in the case of a free motion, can have six independent degrees of freedom in space. The total degree of freedom f (in mechanisms and gear trains in German also referred to as \u2018\u2018Laufgrad\u2019\u2019) relative to the reference body are given by f \u00bc 6nB XnG i\u00bc1 6 fGi\u00f0 \u00de or \u00f03:4\u00de f \u00bc 6 nB nG\u00f0 \u00de \u00fe XnG i\u00bc1 fGi \u00f03:5\u00de with the number of kinematic loops being nL \u00bc nG nB; \u00f03:6\u00de the following results from Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003422_ecce.2016.7855360-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003422_ecce.2016.7855360-Figure5-1.png",
        "caption": "Fig. 5. Cross-section of the FEA models. (a) Model 1. (b) Model 2",
        "texts": [
            " the gear ratio increases, the optimal values of o/tm and hPM/g will increase and decrease, respectively, for getting the maximum effective air gap flux density, viz. for getting the maximum back EMF and torque of the magnetic-geared PM machines. V. FEA RESULTS To validate the analytical results above, two FEA models of the magnetic-geared PM machines, with the same stator and different numbers of the PM pole-pairs and modulation blocks, are built in this section. The schematic and main parameters of the FEA models are shown in Fig. 5 and listed in Table III. Fig. 6 shows the open-circuit magnetic field characteristics of the model 1. There is a obvious modulation magnetic field with 1 pole-pairs. The fundamental magnetic field of the PM in outer air-gap is reduced significantly attribute to the flux leakage. The 14th-order harmonic of the air gap flux density is mainly produced by the fundamental MMF of PM interaction with the 2nd-order harmonic air gap permeance. Due to the even-order winding factor of the model 1 is zero, the 12th- and 14th-order harmonics of the air gap flux density do not work"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002529_s106345411602014x-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002529_s106345411602014x-Figure1-1.png",
        "caption": "Fig. 1. A trolley with two pendulums.",
        "texts": [
            " The idea of this approach is to search the acceleration of a trolley providing optimal solution of the problem instead of the control force applied to the trolley. After this acceleration is determined, one can also determine the horizontal force applied to a trolley. We must note that the idea of determining the control in a polynomial form based on integrodifferential equations method was also suggested by G. V. Kostin and V. V. Saurin [8, 9]. Similar problems are also actively being studied by M. A. Chuev [10]. Consider the horizontal motion of a trolley of mass m carrying two pendulums with masses of m1, m2 and lengths of l1, l2, respectively (Fig. 1). To describe the system motion, we will introduce horizontal movement of the trolley x and pendulum rotation angles \u03c61 and \u03c62. We solve the problem of oscillation damping, which determines the horizontal control force F that, in a given time moves a trolley by distance S under the condition that the mechanical system transits from the initial rest state into the new rest state. To provide this, the following boundary conditions must be met: (2.1) The differential equations of a system motion in a linear setting are as follows: (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001724_ssd.2015.7348203-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001724_ssd.2015.7348203-Figure5-1.png",
        "caption": "Fig. 5. Set of sectors beams",
        "texts": [
            "head(\u03c9) (5) This function takes into consideration the distance to obstacles \u201ddist (v,\u03c9)\u201d (giving preference of traveling longer distances without colliding with obstacles), the speed \u201dspeed(v)\u201d (preferring to travel at faster speeds) and heading \u201d head(\u03c9)\u201d (giving the progression of the system towards the goal). The \u03b1i values indicate the relative weight to be given to each term in the objective function. BCM obtains a divergent radial projection model of the environment based on the sensors\u2019 common position. The projection model is defined by Fig. 5. We can deduce ,from Fig.6, the geometric relations to calculate the angles: d = \u221a x 2 obs + (yobs) 2 \u2212 robs \u03b8obs = arctan ( xobs yobs ) \u03b8 = arcsin ( robs dobs ) \u03c11 = \u03b8obs \u2212 \u03b8 \u03c12 = \u03b8obs + \u03b8 The model is simplified and then a set of possible candidate beams is determined. After that, the best beam is calculated by maximizing the following objective function : f(\u03c11, \u03c12, d) = \u03b1 ( d cos(|\u03b5|) dmax ) \u2212 \u03b2 \u2223\u2223\u2223 \u03b5 \u03c0 \u2223\u2223\u2223 (6) where \u03b1 and \u03b2 are weight constants adjusted by experimentation with \u03b1+\u03b2 = 1 ; d cos(|\u03b5|)/dmax is the projected distance over the goal direction where \u03b5 is the angle between the goal direction and the robot\u2019s current orientation, defined as \u03b5 = \u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 0 if \u03c11 \u2264 \u03c10 \u2264 \u03c12 \u03c10 \u2212 \u03c12 if \u03c12 < \u03c10 \u03c11 \u2212 \u03c10 otherwise (7) where \u03c10 is the goal direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure2-1.png",
        "caption": "Fig. 2. Two Examples of the Sliding Type of Floating Shoe a Lockheed. b Girl@ two-leading-shoe.",
        "texts": [
            " The first group might aptly be described as \u201cpivoted\u201d; a word which would then have to be avoided in relation to the floating types. \u201cNon-floating\u201d is not an attractive title, although adequate. \u201cAnchored\u201d is a possible alternative, but as ships float when at anchor it would be a misnomer if applied to shoes which do not float. \u201cRigid\u201d has been used, but is unsuitable since the shoe may deliberately be given flexibility. The following nomenclature is therefore proposed :- (1) Pivoted shop (Fig. 1). (2) Floating shoes : (a) Sliding (Fig. 2). (b) Articulated (Fig. 3). Other items are defined as follows :- Leading shoes: shoes in which the movement of the Trailing shoes: shoes in which the movement of the * An alphabetical list of references is given in Appendix 11, p. 51. drum over the lining is towards the pivot or abutment. drum over the lining is towards the applied load. Toe and heel: the ends of the lining in relation to the \u201cankle\u201d at the pivot or abutment. Shoe tip : the part of the shoe at which the operating load is applied; that is, the end remote from the pivot or abutment",
            " Because a sliding shoe can centre itself in the drum, the centre of the wear pattern is not fixed, as with the pivoted shoe, by the motion about its pivot, but depends upon the coefficient of friction between the lining and the drum (Fig. 5). The lining has to be mounted on the shoe in a position Effect of change of angle of friction on the position of the centre of pressure and point of maximum wear. which anticipates a certain friction value. If this is exceeded, the lining will wear more at the heel if the shoe is leading or at the toe if it is trailing, and vice-oersa if the friction is less than expected. In the G i r k g 2 L.S. design (Fig. 2b) the shoes are completely reversible, and the wear pattern is centralized by inclining the face at the shoe tip relative to that at the abutment, but with the parallel abutments of the Lockheed design (Fig. 24, the lining has to be moved round the shoe to the best position. For this reason, the leading and trailing shoes of the latter type are not interchangeable; moreover, the wear pattern is considerably displaced for reverse rotation, but these features do not appear to be detrimental. In spite of the shift of pressure distribution with change of friction value, the output of sliding shoes is less sensitive to variations of friction than any of the pivoted types, if the theoretical analysis is to be relied on",
            " The conclusions which can be drawn from the table are that the sliding shoe tends to have a lower output than the pivoted type and to be more stable, but that it is possible to obtain a very similar performance with either type of shoe, by suitable selection of 2 and h. I = 1.15 and K = 0.75. -drae shoetip load\u2018 F = Pivotedshoes . . . . . . 11- atrailing Sliding shoes with parallel abutments . . . 3leading 4trailing 5 leading 6trailing Sliding shoes or Huck type with 21 deg. abutments at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from INTERNAL EXPANDING SHOE The factors of floating shoes of the Huck type (Fig. 3a) and those with sliding abutments inclined to one another (Fig. 2b) cannot be expressed as simple formulae but they are easily obtained by graphical methods. They are plotted in Fig. 6 for a typical shoe, of which the shoe-tip and abutment forces act at 21 deg. to each other, together with the factors of pivoted shoes, and sliding shoes with parailel abutments, with I = 1.15 and k = 0.75 in each case. The relative inclination of the abutments does not affect the value of p at which the leading shoe will sprag, which will occur when cosece = l / k for all types of floating shoe"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure7.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure7.11-1.png",
        "caption": "Fig. 7.11 The diagrammatic figure: processing of solid ground curing",
        "texts": [
            " Since the main principle of SGC is based on the exposure of each layer of the model by means of a lamb though a mask, the time of process is independent on the complexity of each layer. For example, an ultraviolet (UV) emitter will hit the whole printing area. Also, SGC may involve the milling/grinding process, which is definitely special in rapid prototyping field. For the operation, the SGC machine uses the ionic graphic process to generate an electrostatic image with recyclable toner on a reusable glass plate which forms a mask at the end (Fig. 7.11). The operation procedures are illustrated in Fig. 7.12. This mask is placed over the liquid photopolymer. A UV lamp then shines. UV light cannot penetrate the toner pattern, meaning that UV only shines through the unmasked part (the outline) and causes the polymer to harden along the outline. Next, the liquid polymer is removed, while the interspace is covered with molten 7.5 Rapid Prototyping 203 wax again. A cold plate is then placed over the wax, cooling down the wax such that it solidifies"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002657_978-3-319-44087-3_29-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002657_978-3-319-44087-3_29-Figure1-1.png",
        "caption": "Fig. 1 a High pressure compressor blisk, b finite element sector model and c fundamental blade mode shape (1st flap)",
        "texts": [
            " In the current paper the effect of intentional mistuning is addressed with respect to a mitigation of the forced response. Reduced order models based on the subset of nominal system modes (SNM) [7] are employed in which mistuning is quantified by stiffness variations. Aeroelastic coupling effects are considered employing the method of aerodynamic influence coefficients (AIC) which are put into the SNM-model as described in [8]. Mistuning patterns which are derived from genetic algorithm optimizations are exemplarily analyzed for a 29-bladed high pressure compressor blisk rotor (Fig. 1a) with focus on the fundamental blade bending mode. Additional random mistuning is superimposed in order to evaluate the robustness of the forced response reduction. Finally, the effect of aerodynamic mistuning is analysed. Aiming at a preparation of a numerical model for the forced response computation, a finite element sector model is set up in a first step representing the tuned and cyclic symmetric blisk with identical blades (Fig. 1). Subsequently an eigenvalue analysis is carried out in order to gain the relevant information about the basic vibration characteristics and blade mode families. Focusing on the frequency range around the first fundamental blade mode family, a basic SNM-model with just 31 degrees of freedom is derived. This basic modal model can be easily adjusted to arbitrary blade frequency mistuning by means of a stiffness adjustment of every blade, for details please refer to [7, 8]. These models are valid as long as no change of blade mode shapes appears"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002869_978-3-658-14219-3_42-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002869_978-3-658-14219-3_42-Figure8-1.png",
        "caption": "Fig. 8 Von Mises stress on the optimized upright.",
        "texts": [
            " Structural performances of the obtained solution are reported in Tab. 2 and compared with actual components. Tab. 2 Structural performances of the obtained solution and comparison with the actual component. Performance index Value Variation w.r.t. actual component Caliper mass [kg] 1.063 -22 % Upright mass [kg] 1.529 -23 % D1 [mm3] 375 -15 % D2 [mm3] 201 -47% D3 [mm3] 192 -42% A preliminary stress analysis on the obtained shapes has been performed for evaluating the structural integrity, the obtained stress field is depicted in Fig. 7 and Fig. 8. The obtained stress levels are quite acceptable in all the components, obviously the geometry has to be refined and more detailed evaluation is required, maybe with a kind of shape optimization in the critical areas for reducing the stress values at the notches [10, 11]. In any case the undoubtful advantage is that an optimized solution has been obtained at a very early stage of the design process. The obtained solution is a very good starting point for the development of the new components. Conclusions In this paper the lightweight design of the brake caliper and upright of a race car has been dealt with"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002622_ssp.251.49-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002622_ssp.251.49-Figure1-1.png",
        "caption": "Fig. 1. Wire rope",
        "texts": [
            " Wire rope is twisted into strands out of high-quality thin 0.5- 2 mm wires and the strands are twisted into a rope. Contemporary technology operation and safety requirements are becoming increasingly strict. As the service life of the product grows longer and the materials improve, relative load on the equipment increases and therefore, early defection of faults and periodic diagnostics amount to a large part of technology operating costs. Wire rope is type of cable which consists of several strands of metal wires laid (twisted) into a helix (Fig. 1). Wire ropes used in lifting and transportation equipment are among such components. Their defects are specifically regulated; however, the efficiency of detecting the defects of wire rope and especially the number of broken outer layer wires per length unit and the speed of examination remain a significant technical problem [1\u20134]. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002939_eeeic.2016.7555663-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002939_eeeic.2016.7555663-Figure1-1.png",
        "caption": "Fig. 1: Stator and rotor for a 6x4 SRG.",
        "texts": [
            " In this paper simulation results are presented considering the SRG operating with power control and firing angles optimization in order to provide better performance in wind energy systems. The switched reluctance machines have a rotor with absence of windings. The phase windings are concentrated only in stator. Both rotor and stator are constituted by ferromagnetic material and have salient poles. Thus, the machine is suitable for a wide speed range operation. There are many configuration for switched reluctance machines [4]. Fig. 1 shows the laminated packet for a 6x4 switched reluctance machine. As stated previously, phase windings are concentrated only in stator. Thus, for a 6x4 SRG, there are three phases. This windings must be individually energized. The power application for each winding phase is provided by a power converted, as shown in Fig. 2 [4]. Torque production occurs by the tendency of alignment of stator and rotor poles when the respective phase is energized, position of minimum reluctance for the established magnetic circuit (and maximum inductance)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.12-1.png",
        "caption": "Fig. 35.12 Self-aligning ball bearings with cylindrical bore.",
        "texts": [],
        "surrounding_texts": [
            "Self-aligning ball bearings have two rows of balls and a common sphered raceway in the outer ring and this feature gives the bearing its self-aligning property which permits a minor angular displacement of the shaft relative to the housing. These bearings are particularly suitable for applications where misalignment can arise from errors in mounting or shaft deflection. A variety of designs are available with cylindrical and taper bores, with seals and adapter sleeves and extended inner rings. Angular contact ball bearings (Fig. 35.13) In angular contact ball bearings the line of action of the load, at the contacts between balls and raceways, forms an angle with the bearings axis. The inner and outer rings are offset to each other and the bearings are particularly suitable for carrying combined radial and axial loads. The single row bearing is of non-separable design, suitable for high speeds and carries an axial load in one direction only. A bearing is usually arranged so that it can be adjusted against a second bearing. A double row angular contact bearing has similar characteristics to two single bearings arranged back to back. Its width is less than two single bearings and it can carry an axial load in either direction. These bearings are used for very accurate applications such as the shafts in process pumps."
        ]
    },
    {
        "image_filename": "designv11_64_0002947_s1062739115060439-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002947_s1062739115060439-Figure4-1.png",
        "caption": "Fig. 4. Liquid Controller for Induction motor Speed Control System.",
        "texts": [
            " There are different control methods for slip ring winder drives. Resistance control technique [4] is used only in slip-ring induction motor for the reduction of speed which is directly proportional to the power dissipation in the rotor circuit. The drawbacks of this method are wasteful energy at low speed hoisting and torque is not remaining constant. Many mines hoist use liquid resistance controller for speed control of the motor even today. A control layout of the liquid resistance control is shown in the Fig. 4. JOURNAL OF MINING SCIENCE Vol. 51 No. 6 2015 Cascade Speed Control Method [4] is used in some mines in winder drive. Here, two induction motors are coupled mechanically. First motor of the wound rotor type is connected to normal three phase supply and the second motor\u2019s stator gets supply from first one\u2019s rotor. If both motor torques are operated in the same direction the speed is below synchronous and for opposite direction the speed is above synchronous. But this method has poor efficiency because of two motors and also it operates at a poor power factor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001217_00207179.2014.935959-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001217_00207179.2014.935959-Figure1-1.png",
        "caption": "Figure 1. The trajectories of the leader and the 10 followers for 29 s. The red dashed line is the dynamic trajectories of the leader and the small red circle is the original location of the leader. The rest of the lines are the trajectories of the 10 followers. The asterisks are the original locations of the 10 followers.",
        "texts": [
            " Thus, the denominators of two results are the same. The unique differentiation is a refinement of the numerator in Theorem (3.2). In this section, a numerical example is presented to illustrate the above results. In the simulation, the system consists of 10 followers and a leader with dynamics described by Equations (1) and (2), respectively. \u2207x\u03c3 (\u00b7) are different functions in different figures. We describe the dynamics of the leader of followers in a two-dimensional Cartesian coordinate system with g(y) = \u2212y ( 1 \u2212 20 exp(\u2212\u2016y\u20162 0.2 ) ) . Figure 1 shows the trajectories of the leader and the 10 followers, according to models (2) and (1) with \u2207x0\u03c30(\u00b7) = \u2212x0 + 2 and \u2207xi \u03c3i(\u00b7) = \u2212Bxi + C, where B = [0, 1, 1, 1, 2, 2, 0, 2, 1, 0]T, C = [3, 1, 6, 5, 1, 10, 3, 3, 3, 1]T. All of the initial values are generated randomly and the range is [0, 20]. Figure 2 presents the error change law of the 10 followers with time. The coupling factor K = diag{ki0} = diag{0.13, 0.01, 0.17, 0.19, 0.14, 0.15, 0.15, 0.08, 0.13, 0.03}. The coupling matrix W D ow nl oa de d by [ O ak la nd U ni ve rs ity ] at 0 3: 58 0 1 Ja nu ar y 20 15 is as follows: \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000530_s00407-014-0136-6-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000530_s00407-014-0136-6-Figure2-1.png",
        "caption": "Fig. 2 Upper right hand part of Hooke\u2019s September 1685 diagram for a discrete elliptical orbit rotating clockwise under the action of a sequence of impulses toward O that depend linearly on the distance to this center. Some auxiliary lines have been deleted to show its correspondence with Newton\u2019s diagram, Fig. 1, in the De Motu",
        "texts": [
            " toward the sun (Nauenberg 2005a). He then proceeded to apply Newton\u2019s geometrical construction in a novel way, by fixing the relative magnitude of the attractive impulses which Newton had left undetermined in his description of Prop. 6. Hooke assumed that these impulses depended linearly on the distance (Hooke\u2019s law) and obtained graphically the resulting polygonal orbital motion which for this case has its vertices on an ellipse with its center at the center of the impulsive force. His result, shown if Fig. 2 where we have enlarged the upper half of his diagram, is similar to Newton\u2019s de Motu diagram, Fig. 1, except that the body is rotating in the opposite sense. Hooke\u2019s diagram and associated text16 was dated September 1685, a year and a half before the appearance of the first edition of the Principia, but for unexplained reasons, it was left unpublished (Pugliese 1989, 181\u2013205; Nauenberg 2005a, 12). Newton became aware that a discussion and proof of his limits in Prop. 1 was necessary, and in an initial revision of de Motu, he included eleven Lemmas that describe the mathematical foundations\u2014Newton\u2019s geometrical formulation of the calculus\u2014 on which the entire Principia rests"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure1-1.png",
        "caption": "Fig. 1. Forces exerted on a permanent magnet by stator core magnetization and phase currents, respectively. Animation available in [1].",
        "texts": [
            "ndex Terms\u2014Electric machines, finite-element methods18 (FEMs), magnetic fields, torque.19 I. INTRODUCTION20 THIS paper describes means to isolate and quantify elec-21 tromagnetic forces and torques between any two distinct22 parts of an electric motor. This makes it possible, for example,23 to isolate and analyze the torque exerted by the stator core and24 phase coils to a rotor magnet, as illustrated in Fig. 1. A mo-25 tor designer optimizing, for example, the torque waveform can26 identify the interactions that dominate the desired or undesired27 behavior. With an optimization problem, this can, for example,28 lead to a lower dimensional feasible set.29 The approach relies on a well-established decomposition of30 the magnetic field according to its sources. In more detail, once31 the magnetic field of a motor, that is, the pair (H,B), is solved,32 the magnetization and current density of each part of the motor33 can be determined"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001169_1.a32416-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001169_1.a32416-Figure12-1.png",
        "caption": "Fig. 12 One-dimensional analytical model of a wrapped cylindrical model.",
        "texts": [
            " (3), the layer thickness is described with the mechanical properties of the creasing process of the membrane. In this step, the governing equations of the cylindrical model under wrapping around the center hub are derived. The flattening of the cylindrical model is treated based on Aksel\u2019rad\u2019s buckling analysis [8,9]. In the buckling analysis formulated by Aksel\u2019rad, the number of the buckling location is one. On the other hand, the local buckling treated in this paper is induced repeatedly. Thus, we extend Aksel\u2019rad\u2019s buckling analysis to deal with the several buckling locations. Figure 12 presents an analytical model of the wrapped cylindrical model. For the analytical model, we make the following four assumptions. 1) The cylindrical model is one-dimensionallywrapped around the center hub. 2) Local buckling is induced in \u00b7\u00b7 Bn\u22122 and Bn\u22121, and local buckling is about to be induced in Bn. 3) The cylindrical model is separate from the center hub at C. 4) The buckling is repeatedly induced, and the interval is constant. Based on assumption 4, we focus on the membrane element betweenBn\u22121 andC in Fig. 12. In the figure, x,L, \u03b1, and \u03c80 represent the body fixed coordinate of the element, the length of the element, the ratio between the interval of local buckling and the length of the element, and the wrapped angle of the element, respectively. To treat the repeatedly induced local buckling, we introduce a flattening function of the membrane element as w \u03b6; x;\u03b1 a\u03b6 cos 2\u03b8 cos2 \u03b1\u03c0x L (4) wherew, \u03b6, and \u03b8 denote the displacement of the cross section in the normal direction (see Fig. 12b), a dimensionless measure of the flattening of the cross section, and an angular coordinate measured from the neutral line (Fig. 12b), respectively. Using the flattening function, at first, the strain energy of the flattening and that of the circumferential bending are calculated. Then, the governing equation of the wrapped cylindrical model is derived by employing the principle of minimum potential energy. The strain and the displacement of a thin cylindrical shell are given as \u03b5\u03b8 1 a \u2202u \u2202\u03b8 w ; \u03ba\u03b8 1 a2 \u2202u \u2202\u03b8 \u2212 \u22022w \u2202\u03b82 ; \u03bax \u2212 \u22022w \u2202x2 (5) D ow nl oa de d by U N IV E R SI T Y O F C A L IF O R N IA - D A V IS o n Ja nu ar y 25 , 2 01 5 | h ttp :// ar c",
            " We assume the flattening is induced with inextensional deformation, and thus Eq. (5) is equal to 0. Consequently, the displacement in the circumferential direction is obtained as shown next: u \u2212 a\u03b6 2 sin 2\u03b8 cos2 \u03b1\u03c0x L (6) Using Eqs. (4\u20136), the strain energy of the flatteningUflat is calculated as Uflat Et3a 24 1 \u2212 \u03bd2 Z L 0 Z 2\u03c0 0 \u03ba2x \u03ba2\u03b8 2\u03bd\u03bax\u03ba\u03b8 d\u03b8 dx Et3a 24 1 \u2212 \u03bd2 a2\u03b62\u03c04\u03b13 L3 1 2 sin 4\u03c0\u03b1 2\u03c0\u03b1 9L\u03b62 4a2 3\u03c0 2 1 \u03b1 sin 2\u03c0\u03b1 1 8\u03b1 sin 4\u03c0\u03b1 3\u03bd\u03b62\u03c02\u03b1 L \u03c0\u03b1 sin 2\u03c0\u03b1 1 4 sin 4\u03c0\u03b1 (7) where \u03bd is Poisson\u2019s ratio. The \u03b7 coordinate in Fig. 12b of the deformed shape is calculated as \u03b7 a sin \u03b8 w sin \u03b8 u cos \u03b8 a sin \u03b8 \u2212 a\u03b6 sin3 \u03b8 cos2 \u03c0\u03b1x L (8) The second moment of the area of the deformed cross section I is obtained using the coordinate \u03b7 as follows: I \u03c0a3t 1 \u2212 3 2 \u03b6 cos2 \u03c0\u03b1x L 5 8 \u03b62 cos4 \u03c0\u03b1x L (9) Because \u03b6 is a small parameter, the \u03b62 term in Eq. (9), which is shown as the underlined term, is omitted. Because the membrane is wrapped around the cylindrical center hub, the deformed shape of the cylindrical model becomes an arcshaped configuration",
            " In addition, the maximum value of the layer thickness and the membrane thickness, a\u2215t, is about 32 under the finite element method (FEM) and the experimental conditions. Generally, the plastic buckling can be induced when the ratio between the radius of the cylinder and the membrane thickness is less than 50. Thus, the plastic buckling stress is used as the buckling stress. An example of the plastic buckling stress is indicated in Eq. (20), which is in terms of the tangent modulus theory [12] and can evaluate the material nonlinearity simply. \u03c3cr \u2212 Ett\u03ba1 3 1 \u2212 \u03bd2 p (20) where Et and \u03ba1 are the tangent modulus and the curvature of the cross section; see Fig. 12. \u03ba1 is calculable using Eq. (5) as \u03ba1 1 a \u03ba\u03b8 \u03b8 \u2212\u03c0\u22152;x L\u2215\u03b1 1 \u2212 3\u03b6 a (21) Because the condition for local buckling can be described as \u03c3Bn \u03c3cr, using Eqs. (19) and (20), the local buckling condition is obtained as \u2212 Ett 3 1 \u2212 \u03bd2 p 1 \u2212 3\u03b6 a 2Tr 1 \u2212 \u03b6 sin 2 \u03c80\u2215\u03b1 at 1 \u2212 3\u22152 \u03b6 \u2212 T t 0 (22) Using Eqs. (16), (17), and (22), three unknowns \u03b6, \u03c80, and \u03b1 are obtained. Additionally, the interval of the local buckling lb can be calculated using Eq. (22) as lb L \u03b1 r\u03c80 \u03b1 r sin\u22121 a 2\u2212 3\u03b6 4 1\u2212 \u03b6 \u03ba1 3 1\u2212 \u03bd2 p Ett 2 Tr 1 r s (23) Equation (22) is also described as \u2212 1\u22123\u03b6 3 1\u2212\u03bd2 p Et E T Et 1 t\u2215r 2sin2 \u03c80\u2215\u03b1 1\u2212 3\u22152 \u03b6 1\u2212\u03b6 \u2212 T Et 1 l0\u2215r c21 24 1\u22152 0 (24) Therefore, the following three dominant parameters of the local buckling, which mainly determine the condition for local buckling Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002864_0954405416661003-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002864_0954405416661003-Figure1-1.png",
        "caption": "Figure 1. Coordinate systems for derivation of gear.",
        "texts": [
            " Therefore, applying the ease-off topography, the article proposed a novel double-crowned tooth geometry by computer-aided design for spiroid gears manufactured by precision casting process. The goals of developing the geometry of double-crowned gear are to localize the contact and pre-design a controllable function of transmission errors for the helicon gearing. Generally, the hob with the same number of teeth as the spiroid cylindrical pinion is applied to generate the fully conjugated tooth surfaces of the spiroid gears. The coordinate transformation systems are applied for derivation of the gear tooth surfaces as shown in Figure 1. Movable coordinate systems Sh and S2 are rigidly connected to the hob and the gear, respectively; coordinate systems Sp and Sa are the auxiliary coordinate systems that are used for simplification of derivation of coordinate transformation. E is the shortest center distance between axes of rotation of zh and z2. Parameters R2 and R1 determine the outer and inner dimensions, respectively, of the spiroid gear and can be obtained from the conditions of avoidance of undercutting and pointing.14 Angle gm is formed by the axes of the gear and the hob"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.36-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.36-1.png",
        "caption": "Fig. 3.36 Symmetrical portal frame structure: a half model; b free body diagram",
        "texts": [
            "4 Assembly of Elements to Plane Frame Structures 167 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 \u2212 \u2212 qL3A + 6FL2A + 36FI 12 ( AL2 + 3I ) \u2212L ( 6LFA \u2212 7qAL2 \u2212 24qI ) 2 ( 7AL2 + 24I ) \u2212 (\u22127qL5A2 + 72FL4A2 \u2212 3qAIL3 + 900FAIL2 + 72qI2L + 2160FI2 ) L 36 ( AL2 + 3I ) ( 7AL2 + 24I ) \u2212 (qL + 6F)AL2 12 ( AL2 + 3I ) L ( 6FLA + 24qI + 7qAL2 ) 2 ( 7AL2 + 24I ) \u2212L ( 72FL4A2 + 396FAIL2 + 432FI2 + 3qAIL3 \u2212 72qI2L + 7qL5A2 ) 36 ( 7A2L4 + 45AL2I + 72I2 ) \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 . (3.306) (b) In the case of F = 0, the symmetry in regards to the geometry and the load case can be used to create a simplified model under consideration of appropriate symmetry conditions. As can be seen in Fig. 3.36a, only half of the structure needs to bemodeled if at the symmetry line (X = L 2 ) the condition u3X = \u03d53Y = 0 is imposed. The free body diagram of this structure is shown in Fig. 3.36b where now only two finite elements are required to simulate the structure. The elemental stiffness matrices can be taken from Eqs. (3.300) and (3.301) in which the transformation LII = L 2 must be applied. Assembling to the global system of equations and consideration of the boundary condition, i.e. u1X = u1Z = u3X = 0 and \u03d51Y = \u03d53Y = 0, gives the following reduced system of equations: 168 3 Euler\u2013Bernoulli Beams and Frames \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 12I L3 + 2A L 0 + 0 \u2212 6I L2 + 0 0 0 + 0 A L + 96I L3 0 \u2212 24I L2 \u221296I L3 \u2212 6I L2 + 0 0 \u2212 24I L2 4I L + 8I L 24I L2 0 \u221296I L3 24I L2 96I L3 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 u2X u2Z \u03d52Y u3Z \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 0 \u2212qL 4 +qL2 48 \u2212qL 4 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000629_iccse.2015.7250378-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000629_iccse.2015.7250378-Figure8-1.png",
        "caption": "Fig. 8 Productions designed by engineering undergraduate using the TRIZ theory in problem based learning mode",
        "texts": [
            " The climbing robot teaching platform is as shown in Figure 7. Through problem based learning, the engineering undergraduate as the main innovation activities can greatly reflect the engineering undergraduate\u2019s dominant position, but also to maximize the engineering undergraduate\u2019s enthusiasm and initiative. This is an effective way to stimulate engineering undergraduate\u2019s practical innovative ability. This will be in favor of the improvement of engineering undergraduate\u2019s innovation thinking. Figure 8 is productions designed by engineering undergraduate using the TRIZ theory in problem based learning mode. In the actual teaching process, we select this teaching mode according to the character and link teaching content. Its core is to guide engineering undergraduate to apply the TRIZ theory and problem based learning method to raise and stimulate engineering undergraduate\u2019s innovation ability. IV. CONCLUSION As the foundation of knowledge innovation, dissemination and application, education is the cradle of training the spirit of innovation and innovation talent"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001811_0954405415608784-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001811_0954405415608784-Figure4-1.png",
        "caption": "Figure 4. The displacement of a ball bearing due to preload(a), Ball raceway contact after a static load is applied (b).",
        "texts": [
            " For all of the ball bearings, the following equation is employed after adjusting for fatigue life (Zaretsky9) Lna = a1a2a3a4 Qci Qi 10 9 + Qci Qc 10 9 ! 9 10 \u00f05\u00de The multiplicative factor concept employed in equation (2) has been used since the 1960s when the first improvements in bearing steel were made and the role of lubricant films in bearing fatigue endurance was discovered. The lifetime formula used in many bearing applications has substantially exceeded predictions.11 Internal load distribution of ball bearings Figure 4(a) shows the displacement of a ball bearing inner ring relative to the outer ring due to preload Fa. When a ball is compressed by Fa, the distance between the raceway groove curvature and the centre is increased by the contact deformation di (inner ring) and do (outer ring), as shown in Figure 4(b). Under zero load, the centres of the raceway groove curvature radii are separated by a distance BD, where BD= ri + ro D (ri and ro are the radius of the inner and outer raceway groove curvature, respectively, and D is the diameter of the ball). When a centrifugal force, Fc, impacts the ball, Figure 5 assumes that the outer raceway groove curvature centre is fixed in space and that the inner raceway groove curvature centre moves relative to that fixed centre. Moreover, the ball centre shifts by virtue of the dissimilar contact angles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000917_sisy.2015.7325384-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000917_sisy.2015.7325384-Figure1-1.png",
        "caption": "Fig. 1. Modified Denavit-Hartenberg formalism",
        "texts": [
            " A body can be virtual or real: virtual bodies are introduced to describe joints with multiple degrees of freedom such as ball joints or intermediate fixed frames. The body and the segments are numbered increasing from the base body \ud835\udc351 to the terminals. In the sequel, body \ud835\udc35\ud835\udc57 is the successor of \ud835\udc35\ud835\udc4e(\ud835\udc57) and the antecedent of \ud835\udc35\ud835\udc60(\ud835\udc57). Frame \ud835\udc3e\ud835\udc57 , associated with body \ud835\udc35\ud835\udc57 , is given by its origin and an orthonormal basis (\ud835\udc65\ud835\udc57 , \ud835\udc66\ud835\udc57 , \ud835\udc67\ud835\udc57). The transformation between frame \ud835\udc3e\ud835\udc56 and \ud835\udc3e\ud835\udc57 is performed using six geometric parameters \ud835\udefe\ud835\udc57 , \ud835\udc4f\ud835\udc57 , \ud835\udefc\ud835\udc57 , \ud835\udc51\ud835\udc57 , \ud835\udf03\ud835\udc57 , \ud835\udc5f\ud835\udc57 as shown in Fig. 1 and can \u2013 223 \u2013 978-1-4673-9388-1/15/$31.00\u00a92015 IEEE be described by the homogeneous transformation \ud835\udc56\ud835\udc47\ud835\udc57 , see [6]: \ud835\udc56\ud835\udc47\ud835\udc57 = \ud835\udc45(\ud835\udc67, \ud835\udefe\ud835\udc57)\ud835\udc47 (\ud835\udc67, \ud835\udc4f\ud835\udc57)\ud835\udc45(\ud835\udc65, \ud835\udefc\ud835\udc57)\ud835\udc47 (\ud835\udc65, \ud835\udc51\ud835\udc57)\ud835\udc45(\ud835\udc67, \ud835\udf03\ud835\udc57)\ud835\udc47 (\ud835\udc67, \ud835\udc5f\ud835\udc57) = [ \ud835\udc56\ud835\udc34\ud835\udc57 \ud835\udc56\ud835\udc5d\ud835\udc57 01\u00d73 1 ] where \ud835\udc45(\u22c5, \u22c5) and \ud835\udc47 (\u22c5, \u22c5) are the homogeneous transformation of rotation and translation, \ud835\udc56\ud835\udc34\ud835\udc57 defines the (3 \u00d7 3) rotation matrix and \ud835\udc56\ud835\udc5d\ud835\udc57 is the (3 \u00d7 1) vector describing the position of the origin of \ud835\udc3e\ud835\udc57 with respect to \ud835\udc3e\ud835\udc56. In the sequel, a single subscript is used for shorthand if a physical variable is expressed in its own frame such as \ud835\udc57\ud835\udf14\ud835\udc57 = \ud835\udf14\ud835\udc57 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000796_wcica.2014.7053104-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000796_wcica.2014.7053104-Figure1-1.png",
        "caption": "Fig. 1 Robot Coordinate System",
        "texts": [
            " This paper presents an identification method which can overall identify all parameters mentioned above, identification results can be used to compensate for force signal acquired by sensor, to achieve accurate acquire the pure contact force. In order to describe the relationship between six-axis force/torque vector and the robot posture, we need define three coordinate system: the robot base coordinate system {0}, the robot sixth axis coordinate system {6}, sensor coordinate system {S}, the coordinate systems are shown in figure 1. The sensor coordinate system is rotated q around the Zaxis of the robot sixth axis coordinate system according to the right-hand screw rule, 0 6 R describe the robot sixth axis coordinate system respect to the robot base coordinate system, it represent the posture of the robot end; 6 S R describe the sensor coordinate system respect to the robot sixth axis coordinate system. The core idea of the parameters identification as follows: firstly, building a parameter identification model about the parameters need to be identified, the robot posture, and the force measured by the sensor; Then, restructuring the parameters as unknown vector; Finally, using the least squares method to calculate the unknown vector which is the result of parameters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003232_detc2016-59194-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003232_detc2016-59194-Figure6-1.png",
        "caption": "Fig. 6 Transition configurations between the planar 4R mode and the orthogonal Bricard 6R mode of the multi-mode 7R mechanism.",
        "texts": [
            " The transition configuration between two motion modes can be obtained using the set of the equations composed of both set of equations associated with these motion modes. 5.2.1 Transition configurations between the planar 4R mode and the orthogonal Bricard 6R mode Combining the equations associated the planar 4R mode [Eq. (45) or (46)] and the orthogonal Bricard 6R mode [Eq. (47) or (48)] of the multi-mode 7R mechanism, we have \u03b82 = 0 \u03b84 = 0 \u03b86 = 0 \u03b87 = 0 \u00b7 \u00b7 \u00b7 (51) Solving Eq. (51), we obtain the two transition configurations between the planar 4R mode and the orthogonal Bricard 6R mode: Transition configuration 1 (Eq. (52) and Fig. 6(a)) and transition configuration 2 (Eq. (53) and Fig. 6(b)). \u03b82 = 0 \u03b84 = 0 \u03b86 = 0 \u03b87 = 0 \u03b81 = \u03b83 = \u03b85 = \u22122\u03c0/3 (52)  \u03b82 = 0 \u03b84 = 0 \u03b86 = 0 \u03b87 = 0 \u03b81 = \u03b83 = \u03b85 = 2\u03c0/3 (53) 5.2.2 Transition configurations between the planar 4R mode and the plane symmetric 6R mode Combining the equations associated the planar 4R mode [Eq. (45) or (46)] and the plane symmetric 6R mode [Eq. (49) or (50)] of the multi-mode 7R mechanism, we have \u03b82 = 0 \u03b84 = 0 \u03b86 = 0 \u03b81 = \u03c0 \u03b85 = \u03b83 \u00b7 \u00b7 \u00b7 (54) Solving Eq. (54), we obtain the two transition configurations between the planar 4R mode and the plane symmetric 6R mode: Transition configuration 3 (Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003022_icpes.2016.7584071-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003022_icpes.2016.7584071-Figure7-1.png",
        "caption": "Fig. 7. The main machine block made of Simulink blocks.",
        "texts": [
            " This approach is simple to understand due to the fact that all of the electrical behaviour is modelled using power system block-set and all of the control behaviour is modelled using Simulink block-set. The stator RLE model representing the electrical behaviour of the BLDC machine is shown in Fig. 5. The results of this block are the three phase currents which serve as the coupling input for the mechanical model and produce machine torque. The mechanical model, as shown in Fig. 6, consists of two subsystems, namely the main machine and the trapezoid generation block that implements Fig. 1. The main machine block, as shown in Fig. 7, implements (4), (5) and (6) using Simulink blocks such as transfer function, integrator, adder, gain and product. The trapezoid generation block, as shown in Fig. 8, consists of three identical unit trapezoid function blocks. The input to the phase-A block is position, phase-B block is 120\u00b0 shifted position and phase-C block is 240\u00b0 shifted position. The implementation of the unit trapezoid function is shown in Fig. 9. The mechanical model gives the speed and the position as output, and not the hall logic"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure12.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure12.5-1.png",
        "caption": "Fig. 12.5 (a) Basic structure of an artificial knee, which includes a tibial plate as indicated by arrow. (b) Simplified geometry of the tibial plate, with the expected rough surface highlighted in gray",
        "texts": [
            " A competent designer can significantly reduce the amount of thinking time for an optimal process configuration as discussed previously in Sect. 12.3.3. 12.5.1 Milling over a Surface of an Artificial Knee Milling over surfaces of implanted biomedical devices can help define both shape and surface roughness as we previously discussed in Chap. 10. In this section, let\u2019s consider a scenario in which a patient has been arranged for an implantation surgery of an artificial knee joint after a severe car accident. The expected outcome after the surgery is described as Fig. 12.5a. The lower tibial stainless steel plate is confirmed to be covered by a polyethylene sheet, whereas the material for the upper femoral replacement component is only stainless steel. To facilitate the geometric modeling for the following discussion, we may consider the tibial plate as a simplified shape shown in Fig. 12.5b. The side length of the tibial stainless steel plate (Lplate) would be a key geometric parameter to define for the shape and area to be milled. The gray region indicates the surface needed to be milled in order to generate an appropriate surface roughness for better bone cell attachment (as elaborated in Sect. 10.2.3). The milling path includes two stages of slab milling as shown in Fig. 12.6. During milling, we consider that the cutter is always set with the same moving velocity throughout both the milling stages"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.9-1.png",
        "caption": "Fig. 3.9 Revolute joints in a spherical kinematic chain",
        "texts": [
            " In a planar kinematic chain the individual body in general moves with only one rotational and two translational DoF. As a consequence, only the prismatic joint (displacement axis in the motion plane) and the revolute joint (rotation axis perpendicular to the motion plane) of the joints in Table 3.1 remain (Fig. 3.8). In a spherical kinematic chain the individual body in general moves with three rotational degrees of freedom. As a consequence, the spherical motion is comprised of three revolute joints with intersecting axes in the fixed point O of the kinematic chain (Fig. 3.9). In mechanisms and gear trains there is a difference between standard joints (socalled lower kinematic pairs) and complex joints (higher kinematic pairs) (Reuleaux 1875): \u2022 In standard joints, the bodies have surface contact. One distinguishes the following six standard joints (Table 3.3). \u2022 Complex joints have contact along a body line or at a point (Fig. 3.10). Contact between two bodies can occur on two non-physical, spatially fixed surfaces or rather body fixed surface, spur surface or rather pin surface, along the instantaneous axis of rotation or rather screw axis, which indicate the instantaneous motion state"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002716_snpd.2016.7515943-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002716_snpd.2016.7515943-Figure1-1.png",
        "caption": "Figure 1: Flight Parameters of Unmanned Aircraft along the three coordinate.",
        "texts": [
            " +0, ---(1 ) This parameter such defined oE is Elevator angle, oEmin is Elevator angle in trim condition, oe is Elevator angle Compared to amount of trim, oR is Rudder angel, oA is Aileron angle, h is height, <I> is Roll angle, is Pitch angle, 'P is Side angle, P is Rate of Roll angle, Q is Rate of pitch angle, R is Rate of yaw angle, U Longitudinal velocity, W is Vertical velocity, V is Lateral velocity, Vt is the total Rate, a is Attack angle and P is lateral movement angle. Some variables of UA V are defined in the Figure 1. The purpose is design the Autopilot for height, which is capable to control the height of Aircraft by using the elevator height. Equation 1, shows the nonlinear behavior UA V. By linearization the nonlinear Equations around trim flight Cruise condition, nominal linear model for height be made in the Transfer Function or Equation 2. h(s) - 57.3(s - 24.6)(s+ 21) (s+0.008) S,(s) S(Sl + 0.01ls+0.022)(s' + 2.12s+98.4) ---(2) It is assumed that the nominal linear model is a mathematical model, so that it developed based on Autopilot"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000554_9781118774038.ch15-Figure15.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000554_9781118774038.ch15-Figure15.17-1.png",
        "caption": "Figure 15.17 Schematics showing the process fl ow to achieve G/ZnO based sensor. (a)\u00a0Patterning the comb-like Zn electrodes, (b) oxygen plasma bombardment to form ZnO nanowires, (c) coating the graphene oxide sheets on the obtained structure, and (d) photo-catalytic reduction of graphene oxide using ZnO nanowires (Reproduced from [142] with permission from Elsevier).",
        "texts": [
            " Th e sensitivity and selectivity of graphene sensor can be improved by using graphene/metal oxide nanocomposites. Th e enhanced sensitivity of the device is originated by the charge transfer transition of metal oxides to the graphene sheets. Several metal oxides such as, ZnO, SnO 2 , TiO 2 , WO 3 , NiO, Fe 2 O 3 , Cu 2 O, etc., have been used for the development of gas-sensing device [132\u2013149]. Afzali et al. developed graphene-based oxygen sensor using photocatalytic activity of ZnO nanowires [142]. Th e schematic of the designed sensor is shown in Figure 15.17. In the case of ZnO nanowire sensor, the resistance was found to increase with oxygen exposure. In contrast, the rGO/ZnO sensor exhibited the decrease of resistance upon oxygen exposure. Th is may be attributed to the adsorption of oxygen molecules on the surface of graphene sheets resulting in the increasing hole conduction as a p-type donor. Th e sensitivity of the designed sensor was found to be signifi cantly higher than that of the existing graphene-based oxygen sensor [143]. Th e sensitivity of the sensor was found to increase at elevated temperature (200oC)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001001_20140824-6-za-1003.00348-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001001_20140824-6-za-1003.00348-Figure2-1.png",
        "caption": "Fig. 2. Open and closed states of the gripper",
        "texts": [
            " Two linkages, each consisting of a universal joint of the screwdrive and a spline nut, are for achieving omnidirectional bending motion. A third linkage is for achieving rotary motion of the gripper. 978-3-902823-62-5/2014 \u00a9 IFAC 7233 Opening and closing motions of the gripper are attained by wire actuation. Only one side of the jaws can move, and the other side is fixed. The wire for actuation connects to the drive unit through the inside of the DSD mechanism and the rod, and is pulled by the motor. The closed and open states of the gripper are shown in Fig. 2. The built DSD forceps and its controller are shown in Fig. 3. The main specifications of the DSD forceps and the detailed explanation for the mechanism are illustrated in Ishii et al. (2010). As shown in Fig. 2, wire actuation mechanism is employed for opening and closing motions of the jaw. Then, to detect the grasping force of the jaw without using sensor is desired. In order to analyze the wire actuation mechanism, an enlarged model of one side of the jaw was built and a simplified wire actuated forceps model is introduced. Derivation of a new algorithm based on the RFOB is developed for the simplified model. The built enlarged model of the jaw is shown in Fig. 4. A DC motor rotates a motor pulley. The motor pulley and a joint coupling are connected by a wire passing through the inside of a rod"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003134_978-981-10-2875-5_70-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003134_978-981-10-2875-5_70-Figure6-1.png",
        "caption": "Fig. 6 Equivalent mechanism of the tetrahedral element",
        "texts": [
            " According to the reciprocal screw theory the twist of the node B can be derived as ( )T m 0 0 0 0 1 0B/ =S , \u00f05\u00de where m B/S represents a translation along the y-axis, which shows that the node B has a translational DOF with respect to the fixed node A. As similar as the above solving process, it can be known that the node C also has a translational DOF with respect to the node A. On the condition of considering the constraint influence of the second part of the tetrahedral element, the tetrahedral element can be equivalent to the mechanism shown in Fig. 6. It can be easily obtained that the number of the DOFs of the equivalent mechanism is one. According to the analysis in Sect. 2.2 the equivalent mechanism of the minimum composite unit shown in Fig. 2 can also be formulated, as shown in Fig. 7. Assuming that the foldable and deployable strut BG is removed from the equivalent mechanism shown in Fig. 7 and the joint connecting the node B to the node A is selected as the actuated joint, then the nodes C, D, E, F and G will move as the movement of the node B, as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.53-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.53-1.png",
        "caption": "Fig. 3.53 Finite element approximation with a single beam element: a simply supported beam; b cantilevered beam",
        "texts": [
            " \u2022 Calculate based on the appropriate partial differential equation the analytical solution for the deflection uz = uz(x) and compare this result with the finite element solution at x = L. \u2022 Sketch the analytical and finite element solution, i.e. uz = uz(x), in the range 0 \u2264 x \u2264 L. 178 3 Euler\u2013Bernoulli Beams and Frames Fig. 3.52 Cantilevered beam with triangular shaped distributed load 3.28 Finite element approximation with a single beam element Given is a beam with different supports as shown in Fig. 3.53. The bending stiffness EI is constant and the length is equal to L = a + b. The beam is loaded by a single force at location x = a. Derive the finite element solution based on one single beam element and compare the displacements uz(0), uz(L) and uz(a) with the analytical solution. 3.29 Cantilevered beam: moment curvature relationship The cantilevered beam shown in Fig. 3.54 is loaded by a single forceF in the negative Z-direction at its right-hand end. The total length of the beam is 2L and the bending stiffness is EI "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.43-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.43-1.png",
        "caption": "Fig. 10.43 Ac inductor, resonant converter example: (a) resonant tank circuit, (b) inductor current waveform",
        "texts": [
            " Typically, the core loss can be ignored, and the design is driven by the copper loss. The maximum flux density is limited by saturation of the core. Proximity losses are negligible. Although a high-frequency ferrite material can be employed in this application, other materials having higher core losses and greater saturation flux density lead to a physically smaller device. Design of a filter inductor in which the maximum flux density is a specified value is considered in the next chapter. An ac inductor employed in a resonant converter is illustrated in Fig. 10.43. In this application, the high-frequency current variations are large. In consequence, the B(t) \u2212 H(t) loop illustrated in Fig. 10.44 is large. Core loss and proximity loss are usually significant in this application. The maximum flux density is limited by core loss rather than saturation. Both core loss and copper loss must be accounted for in the design of this element, and the peak ac flux density \u0394B is a design variable that is typically chosen to minimize the total loss. A high-frequency material having low core loss, such as ferrite, is normally employed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001842_amm.658.377-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001842_amm.658.377-Figure1-1.png",
        "caption": "Fig. 1 Classic gradient for temperature in a metal wall [6]",
        "texts": [
            " The temperature fall is located under the surface of the metal wall at a very small depth on both surfaces of the wall ( on the entry and exist surfaces of the thermal flow ). The hypothesis using as decisive stresses \u03c30, \u03c445D and \u03c3yt, that have the maximum values on the contact surface, gives no explanation for the origin of the destructions under the contact surface, fact demonstrated when the \u03c40, \u03c4c and \u03c3ED(\u03bb) hypothesis are taken into account. The equivalent stress \u03c3ED(\u03bb) has a high enough value on the contact surface, thus giving a global explanation for the destruction origin points, both on and under the contact surface. In Fig. 1 and 2 [6], the notifications have the follow means: tp.cl = wall temperature, following Fourier\u2019s law, tp.r = the real temperature distribution for the wall, taking into account the thermal Jacq effect, \u2206ti = tpr1 \u2013 tpr2, the thermal fall after Jacq effect, \u2206tcond is the thermic fall in conduction transfer (the classically calculated value for the temperature). The thermal gradient is different in these two methodologies. The importance of this thermal effect for the position of primary destruction point is that the change of slope in the temperature variation very near to the surface indicates a variation of thermal stress in that region"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000392_978-1-84996-432-6_26-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000392_978-1-84996-432-6_26-Figure1-1.png",
        "caption": "Fig. 1. Experimental set-up",
        "texts": [
            " The effect of pneumatic barrier in fluid flow through grinding zone is also observed . In the present work, the grinding wheel is fitted on a three axis CNC milling machine (Make- BFW, India, ModelAkshara VF 30 CNC). The wheel velocity of 20.9m/s has been chosen with a wheel diameter of 200mm. Specification of the wheel used is AA46/55K5V8. A special type calibrated probe, made indigenously, is used to measure air pressure at various points radially outwards from the wheel peripheral surface (R-axis) as shown in Fig. 1. The probe of 1mm outer diameter is set at 0.1mm distance from the middle position of wheel periphery. U-tube inclined manometer with water as manometric fluid and inclination angle of 15o is used to measure the air pressure. A pneumatic nozzle of 4mm internal diameter is placed on wheel face at different polar co-ordinates (r, \u03b8) (Fig. 2). The polar angles (\u03b8) considered are 30o, 45o and 60o. The pneumatic nozzle is placed 10mm away from the middle position of the wheel periphery, so that curve radius, r becomes 110mm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002832_978-981-10-1956-2_5-Figure5.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002832_978-981-10-1956-2_5-Figure5.19-1.png",
        "caption": "Fig. 5.19 Schematic diagram of the three degree of freedom redundant manipulator",
        "texts": [
            " However, as stated before without demonstrating application of the control law on a robotic system this chapter would not take its complete shape, in the next section a 3-DOF manipulator example is considered. A planar 3-DOF robot arm is a particular type of robot manipulator with two prismatic and one revolute joint, moving on a horizontal plane. The two prismatic joints are rigid, whereas the revolute joint is coupled to the end-effectors through an elastic degree of freedom. In addition, all the prismatic joints are actuated [5]. An idealized model of this manipulator is shown in Fig. 5.19. The model consists of a base body, which can translate and rotate freely in the plane, and a mass-less arm at the tip of which the end effector is attached [5]. The base body is connected to the mass-less arm by a linear torsional spring whose neutral position is u \u00bc 0 [28]. The Cartesian position of the base body as well as the angle through which the base body is rotated can be controlled [28]. The variable u measures the deviation of the mass-less arm from the assigned (u = 0) value [28]. Whenever the variable is displaced from zero, it induces a restoring torque \u2212Ku, where K denotes the torsional spring constant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure14.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure14.8-1.png",
        "caption": "Fig. 14.8 Symmetrical load characteristic. a Internal resistance via its voltage. b Current of the internal resistance via its voltage and transfer characteristic. c Current via internal conductivity",
        "texts": [
            " Then, the fundamental expressions obtain the view m\u00f0Q\u00de \u00bc \u00f00 VQ V \u00fe M 1\u00de \u00bc \u00f00 0:5 A ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi A\u00f0A 1\u00de p 1\u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffi A 1 A r ; m \u00bc m\u00f0V\u00de \u00bc \u00f00 V V \u00fe M 1\u00de \u00bc V 1 V ffiffiffiffiffiffiffiffiffiffiffi A 1 A r \u00bc m0\u00f0V\u00de ffiffiffiffiffiffiffiffiffiffiffi A 1 A r ; \u00f014:13\u00de The scale hyperbolic distance H\u00f00:5\u00de \u00bc Ln\u00bdm\u00f00:5\u00de : The inverse expression m\u00f0V\u00de \u00bc m\u00f00:5\u00de\u00bd D; or m0\u00f0V\u00de \u00bc V 1 V \u00bc A 1 A \u00f0D 1\u00de=2 : \u00f014:14\u00de If it is necessary to set any equal deviation D for different A, we find m0(V) using the second member of (14.14) and then, we get the value V by (14.13). The corresponding mapping of V, m is shown in Fig. 14.7 too. 14.3 Deviation from the Maximum Load Power Point 399 14.4 Symmetrical Load Characteristic for the Full Area of the Load Voltage Variation Let us return to Fig. 14.5. Two lines 1 for the corresponding areas V \u2264 VOC, V \u2265 VOC we are replacing by one hyperbola 2 shown in Fig. 14.8a. In this case, initial Eq. (14.2) obtain the following view Ri ROC i 2 1 A2 Vi VOC 2 \u00bc rOCi ROC i 2 \u00bc r2; \u00f014:15\u00de where Vi = VOC \u2212 V is the internal resistance voltage and a value ri OC \u2265 Ri OC. Hereinafter, the subscript \u201c1\u201d for Ri1 and Ii1 is not used. From here, using the equivalent circuit in Fig. 14.4, we obtain the normalized load characteristic equation I\u00f0Vi\u00de AISC \u00bc Vi=rAVOCffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi A2r2 \u00fe Vi=rAVOC 2q : \u00f014:16\u00de The plot of this expression has the typical view in Fig. 14.8b. If to introduce the coordinate system ID 0GS VGS, this curve is close to the typical transfer characteristic ID(VGS) of MOSFET transistors and its approximation by a hyperbolic tangent [6]. The values ID C, VGS C correspond to the cusp of this curve. 400 14 Quasi-resonant Voltage Converter with Self-Limitation \u2026 Also, it is possible to note that expression (14.16) is the particular case of Rapp\u2019s model of a solid-state microwave power amplifier [8, 19]. If to express the current I through the internal conductivity Yi = 1/Ri, yi OC = 1/ri OC, the circle equation turns out Yi yOCi 2 \u00fe I AISC 2 \u00bc 1: \u00f014:17\u00de The plot of this circle is shown in Fig. 14.8c. 14.5 Asymmetrical Load Characteristics We may introduce asymmetrical characteristics of a common view too. For example, a transfer characteristic of MOSFET transistor may considerably be differing from a symmetric curve. Case 1 We consider a new position of the known hyperbola in Fig. 14.9a. Let the asymptotes 0y, 0x form a rectangular coordinate system y0x, which turned on an angle \u03b1 concerning the initial system Ri 0 Vi. Therefore, the equation of the hyperbola has the view 14.4 Symmetrical Load Characteristic \u2026 401 On the other hand, for the coordinates of point M, the following orthogonal transformation is known [10] x y \" # \u00bc cos a sin k sin a cos a \" # Vi Ri \" # : \u00f014:19\u00de Then, expression (14"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure7-1.png",
        "caption": "Fig. 7 Constrained relationships of tolerance models a structure of tolerance models, b logic structure of tolerance models",
        "texts": [
            " Except ribs of VV, the accuracy requirements of untagged dimension of blocks and brackets should be executed standard as ISO2768-1, which is \u2018\u2018General tolerance Part 1, Dimension and Angle tolerance for untagged components\u2019\u2019, and class \u2018\u2018f\u2019\u2019 among them are selected as executive standard as Tables 1 and 2. The accuracy requirements of untagged geometric tolerance of blocks and brackets should be executed standard as ISO2768-2, which is \u2018\u2018General tolerance Part 2, Geometric tolerance for untagged components\u2019\u2019, and class \u2018\u2018H\u2019\u2019 among them are selected as executive standard as Tables 3 and 4. Creation of Tolerance Model The dimension feature and constraint for tolerance model can be created in 3D as closed dimension chain according to design basis and reference for installation [7]. Figure 7 reveals structure tolerance models and the relation of logical network in different components of this local IWS. Figure 7a, b are equivalent expression, and every \u2018\u2018?\u2019\u2019 nodes in both patterns could be unfolded into new logical sub-network, so the network tolerance model is a large and complex model. Every relationships between different components are defined and included in this model. In this model, the assembly gaps, including toroidal, poloidal and radial direction gaps, on every plates of every blocks should be defined and named. Such as, the gaps between VV left rib and first block (including ten plates) in toroidal direction are named V-1-L-G-1 to V-1-L-G-10 respectively; And the gaps between second block (including ten plates) and third block (including ten plates) in toroidal direction are named 2-3-T-G-1 to 2-3-T-G-10 respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.2-1.png",
        "caption": "Figure 9.2 Cross section of a 20-pole, solid-pole shoe, three-phase machine with one slot per pole and phase q= 1.",
        "texts": [
            " Adjusting machine inductances can be accomplished by appropriately shaping the electrical steel components. Pole shoes can be mounted on the magnets to produce sinusoidal air-gap flux densities. At the same time, they protect the magnets against both electric and magnetic stresses. Moreover, if appropriately shaped, the pole shoes can protect the brittle sintered magnet material from mechanical damage during manufacturing assembly. Although PMs can tolerate significant pressure, they have very low tensile strength. A possible solid-steel rotor configuration with pole shoes is illustrated in Figure 9.2. The configuration can be used to implement a multipole, low-speed machine. The solid pole shoes, which should be shaped to produce a sinusoidal flux density distribution in the air gap, function also as weak damper windings. Therefore, smooth and quiet operation is achieved. The relative permeability of current hard PM materials is approximately 1, the same as with air, so the effective air gap of a PMSM is relatively large. Because of the large air gap, d-axis armature reaction effects remain low, and the harmonics resulting from the small number slots per pole and phase in the stator do not produce significant torque ripple, a fact that gives the machine certain special characteristics"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003857_0954406215589843-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003857_0954406215589843-Figure1-1.png",
        "caption": "Figure 1. Concept of compliant tilting pad air bearing.",
        "texts": [
            " These simulations are computationally intensive, but are capable of representing important non-linear behaviour of the system of which some examples are shown. The paper ends with some concluding remarks. Note that the models and simulation programs in this paper have previously been presented on the ENOC 2014 conference.15 However, the necessary constraint on length of that contribution limited the amount of details that could be given. This paper presents all details of the models and the simulation programs. Figure 1 shows a schematic planar view of the different bodies of the CTPAB concept and their connections. In this section, a dynamic model of these concepts will be developed, based on the following additional assumptions: . The bearing has three or more tilting pads; . The pads do not influence each other directly, there is no dynamical coupling other than the rotor; . The geometry of the pad inner surface is a cylinder segment; . The pad and the bearing housing are rigid bodies, so if the bearing housing does not move, the motion of the pads is fully determined by deformation of the compliant support; at UNIV OF CONNECTICUT on June 4, 2015pic"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure12-1.png",
        "caption": "Figure 12. Characteristic directions and deformations of a spring: (a) Spring A0B0 in the static equilibrium position and the deformed spring A1B1, and (b) deformed spring A1B1.",
        "texts": [
            " Based on equations (60) and (61), the deflection corresponds to the difference between the length of the spring l and the length of the spring in the static equilibrium position lst \u00bc l lst \u00f063\u00de The following analysis is to show how one can find the deflection of a linear spring in a system that at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from performs small oscillations in an approximate, but very convenient way. To that end, a spring A0B0 in the static equilibrium position is considered (Figure 12(a)). Its length in this position is labelled by lst. When further deformed, the spring takes the position A1B1, when its length is l and both ending points change their position as labelled by the vectors r1 and r2. To determine based on equation (63), the deformations in two characteristic directions are considered. The first one is the t-direction, which is collinear with the direction of the spring in the static equilibrium position. When A1B1 4A0B0 one holds t> 0, and when A1B1 5A0B0, one holds t< 0. The second direction of interest is the s-direction, which is orthogonal to the t-direction, as shown in Figure 12(a) and shown in a simplified version in Figure 12(b). By introducing the unit vector e1 that is collinear with A0B0 or A1B 0, and with the direction that indicates extension, one follows t \u00bc r2 r1\u00f0 \u00de e1, s2 \u00bc r2 r1j j2 t2 \u00f064\u00de Based on Figure 12(b), the deflection can now be expressed as t, s\u00f0 \u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi lst \u00fe t\u00f0 \u00de 2 \u00fes2 q lst \u00f065\u00de The deflection is the function of two variables t and s and can be developed into the Maclaurin series. Using the known fact that the potential energy of oscillatory systems that perform small (linear) oscillations around the equilibrium position is a quadratic form of the generalized coordinate,12 expression (65) is truncated to quadratic order and takes the form t, s\u00f0 \u00de t\u00fe 1 2lst s2 \u00f066\u00de This expression implies that one needs to determine the deflection in the t-direction and the s-direction, but the question of importance here is the order of truncation of the series as well as the order of the expressions of t and s in terms of the generalized coordinate q"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000699_j.ymssp.2015.09.039-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000699_j.ymssp.2015.09.039-Figure7-1.png",
        "caption": "Fig. 7. Optimized acoustic filter and corresponding coefficient; (a) filter cavity, (b) 3D model of acoustic filter, (c) acoustic transmission coefficient, (d) acoustic transmission loss.",
        "texts": [
            " The proposed objective function can be written as OFMinimization \u00bc k1 OF1 k2OF2\u00fek3OF3; \u00f0k1;2;3A \u00bd1;2;3\u2026N \u00de OF1 \u00bc Xf h \u00bc Lower Limit 1 f l \u00bc 1 T\u03c0 ; \u00f0Minimization\u00de OF2 \u00bc Xf h \u00bc Higher Limit f l \u00bc Lower Limit T\u03c0 ; \u00f0Maximization\u00de OF3 \u00bc Xf h \u00bc Duct Cutoff Frequency f l \u00bc Higher Limit\u00fe1 T\u03c0 ; \u00f0Minimization\u00de 8>>>>>>>< >>>>>>>: \u00f024\u00de Here, T\u03c0 \u00bc at , and fl, fh are the lower and higher limit of frequency, respectively. The constants k1; k2 and k3 are the weights assigned to each sub-objective function. These weights have been found from numerical experimentation. In the present case the value of k1; k2 and k3 are set to 1, 3 and 2, respectively. Using the generated dimensions from the optimization process, a 3D model has been prepared which represent the acoustic cavity of the filter using SOLID WORKS s and has been shown in Fig. 7(a). The optimized filter coefficient has been shown in Fig. 7(c). Before fabricating the filter for experimental validation, its acoustic behavior has been evaluated using finite element method (FEM). For this purpose, the acoustic cavity, shown in Fig. 7(a), has been analyzed using a commercially available FEM based software, namely ANSYS s . For understanding the acoustical characteristics of the filter, the acoustic transmission loss (TL) across the filter has been estimated numerically. The potential of FEM in estimating acoustic TL using ANSYS s has been discussed in subsequent chapters. The generated numerical results such as acoustic TL and the corresponding sound power transmission coefficient have been compared with analytical results in Fig. 7(d). From the numerical and the analytical results, it is worth noticing that the performance of the optimized filter is very close to required performance. The designed filter has been manufactured using rapid prototyping machine (FORTUS400mcs ) and has been shown in Fig. 8(a). The acoustic properties of the filter have been measured using B&K s transmission loss tube setup and the TL curve has been compared with the analytical and those from the numerical results (see Fig. 8(b)). A good agreement between all three technique can be observed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure12.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure12.6-1.png",
        "caption": "Fig. 12.6 The milling path and expected surface profiles during the two stages of the surface machining process for defining a surface roughness level",
        "texts": [
            " To facilitate the geometric modeling for the following discussion, we may consider the tibial plate as a simplified shape shown in Fig. 12.5b. The side length of the tibial stainless steel plate (Lplate) would be a key geometric parameter to define for the shape and area to be milled. The gray region indicates the surface needed to be milled in order to generate an appropriate surface roughness for better bone cell attachment (as elaborated in Sect. 10.2.3). The milling path includes two stages of slab milling as shown in Fig. 12.6. During milling, we consider that the cutter is always set with the same moving velocity throughout both the milling stages. The direction of milling paths on the material surface in either stage is along the same 346 12 Process Design Optimization direction. After the first milling stage, a series of grooves are expected to be generated over the surface as shown in the left inset of Fig. 12.6. Between the two stages, the milling head would move from the end position of stage 1 to the start position of stage 2 which will also change its orientation. The second-stage milling is similar to the first stage, except that the direction of \u201cgrooves\u201d generated is perpendicular to those from the first stage. The resultant surface profile is illustrated in the right inset of Fig. 12.6. As a more practical consideration, we would also take other factors such as tooling cost and labor wages (mentioned in Sect. 12.3.1) into account. The processing time for setting up/finalization of the machine process (Tset) is assumed to be 2 min, and the time for reorientation of the milling head between the two milling stages (Tchange) is set as 10 s. For the other costs, the machine cost (M ) is $0.5/min with an overhead ratio (OHm) of 0.1; the labor rate (W ) is $3/min with an overhead ratio (OHop) of 0",
            " For the moving distance of either stage, we could first consider the total area of the process area, which is roughly equal to the plate area subtracted by the rod area (\u00bc (8/9)Lplate 2). The total machined \u201clength\u201d 12.5 Demonstrated Examples 347 (Lmachine) should be around such machined area divided by the mill cutting height, i.e., Lmachine (8/9)Lplate 2/Hmill. The area is machined by multiple parallel lines of milling, and the number of lines is roughly about Lplate/Hmill. As previously discussed in Sect. 9.2.2, we need to add allowance distances on both sides of each milling line. Yet, for the area above and below the middle rod as shown in Fig. 12.6 (left), the middle is blocking the milling paths, and the blocking distance on either side of the rod isDmill/2, compensating the allowance distance of those milling paths. Therefore, the additional moving distance caused by the allowance would be (2/3) Dmill. The total distance of moving path Lcut during stage 1 is Lcut 8Lplate2 9Hmill \u00fe 2DmillLplate 3Hmill \u00bc 2Lplate 9Hmill 4Lplate \u00fe 3Dmill \u00f012:39\u00de Likewise, there should be a similar total distance for a moving path during stage 2. Besides, the moving distance between the two stages (Ltravel) is rather straightforward, equal to Lplate/3 + Hmill"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000218_978-3-319-52219-7-Figure3.18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000218_978-3-319-52219-7-Figure3.18-1.png",
        "caption": "Fig. 3.18 (a) Top view of optical table for horizontal alignment. A plane mirror is put where the reflected beam 1 and 2 cross each other. (b) Side view of optical table for horizontal alignment. An angular deviation of the reflected beam 3 is related to a rotation of the optical axis about y-axis",
        "texts": [
            " Steps 2\u20135 are repeated till the both the reflected beams are parallel to the surface of the optical table, thus bringing the optical to the plane of the incident beams. Next, in order to make the optical axis parallel to incident beams, the steps for the horizontal alignment procedure are: 1. A plane mirror, preferably mounted on a kinematic mount, is placed at the intersection of the two reflected beams. This is not necessarily the focal point since the optical axis may not be parallel to the incident beams yet. 2. The plane mirror is then tilted to send one of the beams to a point on the OAPM such that the reflected beam (beam 3 in Fig. 3.18a, b) is above and in-between the incident beams. To make sure the plane mirror is at the meeting point of the two beams, a thin paper can be held at the surface of the mirror to see the convergence of the two beams. 3. Next, the reflected beam 3 is checked to see if it is parallel to the optical table. It can be done by moving an iris or a fixed aperture on the table with a fixed height. And like before, a Magnetic Beam Height Measurement Tool with alignment holes can be used as well. 4. If the reflected beam 3 is propagating upwards, we need to rotate the OAPM about the y-axis in the direction of the smaller angle of incidence of beam 1 and 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000218_978-3-319-52219-7-Figure5.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000218_978-3-319-52219-7-Figure5.7-1.png",
        "caption": "Fig. 5.7 Liquid Crystal (LC) Metamaterial Absorber (MMA) Spatial Light Modulator (SLM) for THz Applications. (a) 3D cross-section schematic of the MMA array covered with LC. (b) Picture of the MMA SLM device coated with LC. (c) Close-up of MMA unit cells [30]",
        "texts": [
            " Biasing the entire substrate precludes it from being used on systems-on-chip (SoC) applications, a main attraction of solidstate SLMs. The switching speeds are slow due to the large associated capacitance of the substrate and the switching voltage is high related to the breakdown voltage of the substrate. Although advances have been made in liquid crystal on silicon (LCOS) spatial light modulator technology [13], there applications in terahertz have been limited. More recently, metamaterial absorbers embedded in liquid crystals were demonstrated in reflection geometry as terahertz spatial light modulators [30], Fig. 5.7. As shown in Fig. 5.7a, the liquid crystal (LC) forms the dielectric in the split gap of the resonator. By applying a bias voltage and thus electric field across the LC, the polarization of the LC is changed which in turn changes the resonant frequency of the absorber, resulting in voltage controlled modulation. This work showed the viability of using a liquid crystal with metamaterial absorbers with results of 75% modulation depth at 3.76 THz. Although promising, the use of 15 V switching voltage, modulation speeds of only 1 kHz and 70% signal absorbtion are significant drawbacks of the technology to make it a serious contender for THz SLM"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000825_race.2015.7097251-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000825_race.2015.7097251-Figure1-1.png",
        "caption": "Fig. 1. Aeroelastic system",
        "texts": [
            " Neural networks and fuzzy logic systems have gained importance to act as good estimators due to their nonlinear approximations. In the present work, LQR control methodology with a neural based estimator is employed to suppress the aeroelastic responses of an airfoil-flap system. Then its performance will be compared with conventional LQR outcomes. II. MATHEMATICAL DESCRIPTION A two-dimensional airfoil model with flap constraint as third degree of freedom is considered in this problem as shown in Fig. 1. ISBN: 978\u221281\u2212925974\u22123\u22120 From the elastic axis, mid-chord and mass centre is at a distance of ab and x\u03b1b respectively, the values are positive when measured aft of airfoil. The distance c.b and x\u03b2.b are from control surface hinge and mass centre of flap or control surface to the mid-chord. The plunge h, pitch \u03b1, and trailing edge control surface \u03b2 are the three degrees of freedom and the equations of motion of the aeroelastic model are given by [9]: )}({)}(]{[)}(]{[ tFtyKtyD =+ (1) where [ ]TtTtMtLtF )()()()( \u2212= with L(t), M(t), T(t), [D], and [K] denote the aerodynamic lift, the pitching moment, flap torque, mass matrix and stiffness matrix respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure9-1.png",
        "caption": "Fig. 9 Example multibody model of squatting with external weights",
        "texts": [
            " Given the geometries of the body segments, density can be specified for soft tissues, fat, and bones to estimate inertial properties. In addition to body segments, interaction objects should also be created as necessary. For instance, for a squat exercise with external weights, the weights can either be modeled as force components, if grip contact forces were measured, or the whole weight system can be integrated into the multibody model to provide resistance during exercise. An example multibody model of squatting is presented in Fig. 9. In the model, the body segments are all represented as rigid, except the tibia that is modeled as a flexible body to analyze internal strains. The foot-ground contact was simplified to bushing elements, as the subject stands firmly on the ground during the whole exercise not moving his feet with respect to the ground. The external weights were modeled as rigid objects, where inertia properties are determined based on density of the material and geometry of the CAD model of the weights and dumbbell bar"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure7-1.png",
        "caption": "Figure 7. Velocity-based safety device.",
        "texts": [
            " The robot can move by controlling the two motors on the basis of the force sensor signals and the encoder signals. In order to lock Shaft C in clockwise and counterclockwise directions, each drive unit has two velocity-based mechanical safety devices (that is, one velocity-based safety device for locking in the clockwise direction and another velocity-based safety device for locking in the counterclockwise direction). A. Velocity-based Safety Device In this section, we review the structure and mechanism of the velocity-based mechanical safety device briefly (see [13] for more information). Fig. 7 shows the structure of the safety device. Gear A, Plate B and Ratchet Wheel C are attached to Shaft C. Plate C has inner teeth. Each Claw B is attached to Plate B by Pin B and positioned as shown in Fig. 8. Guide Bar B is attached to each Claw B. Three Guide Bars B are respectively inserted into three Guide Holes A of Plate A. Plate A has ratchet teeth. Shaft C rotates Plate A via Torsion Spring. Three Claws B normally rotate together with Shaft C. One end of Linear Spring C is connected to Pin C1 attached to Plate C, and another end is connected to Pin C2 of Frame C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.11-1.png",
        "caption": "FIGURE 9.11",
        "texts": [
            " In other words, the helical buckling solutions at load range of 520 N \u2212 940 N are invalid solutions. Obviously the helical buckling load is higher than the linear buckling load. The DQ results are almost exactly the same as the DSC results presented in Ref. [3], the formulations are validated mutually. Variations of Wn Obtained at Various Applied Axial Loads 1879.3 BUCkLING Of INCLINED CIRCULAR CYLINDER-IN-CYLINDER The postbuckling configurations, obtained by the DQM with small disturbance, are shown in Fig. 9.11. There are two plots in the figure representing the same postbuckling configuration but viewing at different directions. It is seen that the buckling mode is lateral and not helical. At P = 1200 N, the maximum rotation is only 0.9479 radians, i.e., \u03b8 \u03c0= 0.30max for the lateral buckling shown in Fig. 9.11. The postbuckling configurations, obtained by the DQM with large disturbance, are shown in Fig. 9.12. There are also two plots in the figure representing the same postbuckling configuration but viewing at different directions. It is seen that the buckling mode is helical and not lateral. At P = 1200 N, the maximum rotation is 13.532 radians, i.e., \u03b8 \u03c0= 4.31max for helical buckling shown in Fig. 9.12. umax=0.30\u03c0 umax=4.31\u03c0 P\u2212umax Curves with Small and Large Initial Disturbances Postbuckling Configuration (Small Disturbance, P = 1200 N, a = 45\u00b0) 188 CHAPTER 9 GEOMETRIC NONLINEAR ANALYSIS The results obtained by the DQM, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000409_978-3-642-28572-1_38-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000409_978-3-642-28572-1_38-Figure2-1.png",
        "caption": "Fig. 2 Schematics of the proposed approach.",
        "texts": [
            " The analytical description of our tactile exploration approach has been published [7]. Here, we experimentally explore the ability of this approach to tactilely map an unknown structured surface, such as an underwater oil-spilling pipe. This approach requires a manipulator mounted on a stationary base and provided with joint encoders. When exploring an underwater site, the base is the ROV itself, which can lie on the ocean floor or anchor itself to an existing fixed structure such as a pipeline (Figure 2). Extension of the method to account for base movement is relatively straight-forward [18]. Once the base is fixed, the manipulator autonomously moves to explore the environment. The objective is to tactilely create a map of this environment within a given accuracy, by touching the surface with the manipulator\u2019s tip and choosing the movements of the manipulator to minimize the exploration time. The required accuracy is given as a parameter representing the smallest feature to be found by the exploration"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003462_icacdot.2016.7877735-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003462_icacdot.2016.7877735-Figure1-1.png",
        "caption": "Fig. 1 Investigation of Lyapunov stability",
        "texts": [
            " The simulations are performed by varying the controller gains and by adding noise to the plant (shown in FigA, the noise is shown to be present in the friction B of the equation (24) because much of the noise is present in the form of varying friction in the packing gland of the pneumatic globe control valve). Finally, the results are compared to see the efficiencies of the proposed controller and the individual controllers for the above mentioned conditions. Fig.8 Model of the pneumatic control valve with the PI controller Fig.9 Model of the pneumatic control valve with the MRAC A. Simulation with best tuned controller gains without noise The step response without the introduction of the noise in the plant (i.e. ideal condition) for the controller gains given in Table.2 is shown in Fig. 10 and Fig. 1 I. conventional PI controlled response show a lag in the output and the conventional MRAC controlled output show an initial lag and oscillation in the output. Fig. 11 shows that the proposed controller is stable from the very beginning of the simulation, while the conventional MRAC stabilizes after sometime. B. Simulation with best tuned controller gains with noise The response of the plant for a pulse input with the noise introduced in the system is shown in Fig.12 for the controller gains given in Table"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002821_2168-9873.1000146-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002821_2168-9873.1000146-Figure1-1.png",
        "caption": "Figure 1: Co-ordinate system of the bearing configuration.",
        "texts": [
            " Non-recessed multilobe journal bearings give superior performance to recessed or pocketed bearings in addition to their relative ease in manufacturing. Since no work on this topic could be found in the literature, the author studied the type of configuration where the lubricant is supplied at a constant pressure with an attempt to find out a better new configuration. Theory The governing equation is the Reynolds equation is a partial differential equation governing the pressure distribution of an incompressible and isoviscous fluid was first derived by Osborne Reynolds [17] in two dimensions for an incompressible fluid (Figure 1). It can be written in dimensionless form as [18] 23 2 3 ( ) ( / ) 22 p p h hh D L h z \u03bb \u03b8 \u03b8 \u03b8 \u03c4 \u2202 \u2202 \u2202 \u2202 \u2202 + = \u039b + \u039b \u2202 \u2202 \u2202 \u2202\u2202 (1) Under steady state condition equation (1) can be reduce to 2 0 0 2 2 23 32 0 2 0 0 0 0 03 ( / ) 0 p h h h h D L h p p z\u03b8 \u03b8\u03b8 \u03b8 \u2212 \u2202 \u2202 \u2202 + + \u039b = \u2202 \u2202\u2202 \u2202 \u2202 \u2202 \u2202 (2) For axially grooved journal bearings, the boundary conditions are 0 1p = in the groove, 0 0p = at the bearing ends and the pressure is set equal to 0 when the pressure falls below zero. (3) Swift -Strieber boundary condition was applied at the cavitation boundary The equation (2) is solved using Gausss-Siedel method with successive over-relaxation technique "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure4.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure4.9-1.png",
        "caption": "Fig. 4.9 Formulation of the constitutive law based on a classical stress\u2013strain and b generalizedstress\u2013generalized-strain relations",
        "texts": [
            " Similar to the end of Sect. 3.2.2, it is more advantageous in a general approach to work with the generalized stresses s = [ My, \u2212Qz ]T and generalized strains e = [ d\u03c6y dx , \u03c6y + duz dx ]T = [ \u03bay, \u03b3xz ]T since these quantities do not depend on the ver- tical coordinate z. Classical stress and strain values are changing along the vertical coordinate. The representation of the constitutive relationship based on the classical stress\u2013strain quantities and the corresponding generalized quantities is represented in Fig. 4.9. Let us mention at the end of this section that for bending in the x\u2013y plane slightlymodified equations occur compared to Tables4.3 and 4.4. The corresponding equations for bending in the x\u2013y plane with shear contribution are summarized in Table4.6. 4.2 Derivation of the Governing Differential Equation 201 Fig. 4.8 Comparison of the analytical solutions for the Bernoulli and Timoshenko beam (ks = 5/6) for different loading and boundary conditions: a Cantilevered beam with end load; b cantilevered beam with distributed load; c simply supported beam with point load (a) (b) (c) 202 4 Timoshenko Beams 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001655_ecc.2015.7331089-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001655_ecc.2015.7331089-Figure1-1.png",
        "caption": "Fig. 1. The Qball-X4 quadrotor UAV [15]",
        "texts": [
            " One of the main advantage of mixing adaptive control is that, it gives opportunity to analyze the stability of the system by using LTI tools. Mixing adaptive control is described in [12] with details for continuous time, and extended to discrete time in [14]. In this study, we design a mixing adaptive controller for attitude dynamics of the quadrotor. Our goal is to perform lateral tracking of reference trajectories under inertial uncertainties. In this section, we derive equations of motion to be used in our control design later, considering the Qball-X4 (Fig. 1) as the benchmark system. Quadrotors consist of four rotors on a rigid frame [16]. They work as two pairs of propellers. These pairs rotate in opposite directions. Quadrotor performs its motion with the change of speed in these rotors. The whole quadrotor motion dynamics is a 6 DOF system defined by the pitch, roll, yaw angles (\u03b8 ,\u03c6 ,\u03c8 respectively) and the global position [xc,yc,zc] T of the frame. The corresponding nonlinear system equations are derived in [17] using Newton-Euler methods as follows; 978-3-9524269-3-7 \u00a92015 EUCA 3597 \u03b8\u0308 = l(T1\u2212T2) Jx + (Jy\u2212Jz)\u03c8\u0307\u03c6\u0307 Jx \u2212d\u03b8 \u03b8\u0307 , \u03c6\u0308 = l(T3\u2212T4) Jy + (Jz\u2212Jx)\u03c8\u0307\u03b8\u0307 Jy \u2212d\u03c6 \u03c6\u0307 , \u03c8\u0308 = K\u03c8 (T1+T2\u2212T3\u2212T4) Jz + (Jx\u2212Jy)\u03c8\u0307\u03c6\u0307 Jz \u2212d\u03c6 \u03b8\u0307 , x\u0308c = (T1+T2+T3+T4)(sin\u03c8sin\u03c6+cos\u03c6sin\u03b8cos\u03c8) m \u2212dxc x\u0307c, y\u0308c = (T1+T2+T3+T4)(\u2212sin\u03c6cos\u03c8+cos\u03c6sin\u03b8sin\u03c8) m \u2212dyc y\u0307c, z\u0308c = (T1+T2+T3+T4)(cos\u03c6cos\u03b8) m \u2212g\u2212dzc z\u0307c, (1) where T1,T2,T3,T4 are thrust forces; Jx,Jy,Jz are rotational inertia matrices; l is the distance between the center of gravity and the propeller; d\u03c6 ,d\u03c8 ,d\u03b8 ,dxc ,dyc ,dzc are drag coefficients, and K\u03c8 is thrust-to-moment gain"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002252_1754337115577029-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002252_1754337115577029-Figure1-1.png",
        "caption": "Figure 1. Schematic presentation of the JA:Ped3 pedal system showing the placement of the strain gauges (four for each direction): ap \u2013 on each wide inside of the vertical beams (4); ml \u2013 on each narrow inside of the vertical beams (4); and vert \u2013 on the upside and downside of the lower longitudinal beam as shown in Figure (2), same configuration on the diagonal side (2).",
        "texts": [
            " Each pedal is instrumented with 12 strain gauges, with four in a full bridge for each direction. The signals in each direction are linearly proportional to the forces applied and can measure positive and negative values. A unique feature is the decoupling of the power flow, which ensures that force shunts in the measuring device do not result in errors in the acquired forces. Furthermore, the pedal is designed to mount on any pedal crank and measure force in three dimensions: ap, ml and vert directions (Figure 1). Each pedal is equipped with a rotary potentiometer, mounted on the pedal axle for measuring the orientation of the pedal with respect to the orientation of the crank (pedal angle (PA)). This PA facilitates the calculation of tangential and radial forces without an additional crank angle measuring device. The system is calibrated (linearity above 98% in all conditions \u2013 Figure 2) resulting in constant calibration factors for each pedal and each direction and does not require an additional calibration prior to each measurement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002896_ccdc.2016.7531811-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002896_ccdc.2016.7531811-Figure1-1.png",
        "caption": "Fig 1. The process flow diagram",
        "texts": [
            " We give the nonlinear model for control system of the CSTR in this paper, and linearize the *This work was supported in part by the National Natural Science Foundation of China (60804005), and by the Natural Science Foundation of Shan dong Province (ZR2011FQ006), by the Natural Science Foundation of Qingdao City (12-1-4-3-(17)-jch. model by using dynamic feedback linearization method [5-6], then, the output can track to the desired output which has been given by using the solving method of optimal tracking problem 2 THE CSTR MODEL ANALYSIS This paper uses the first order irreversible reaction of non-isothermal CSTR system as the research object, the process flow diagram as shown below in figure 1. When we establishment the model, we have done the following realistic assumptions. 1): The mixing is complete; 2): The volume flow rate of the discharged material is equal to the volume flow rate of the inflow. Thus, the reactor can be treated as a whole then modeling the CSTR system. The properties can be represented by the following continuous time nonlinear differential equations: 0( ) exp( )A AF A A dC F EC C K C dt V RT = \u2212 \u2212 \u2212 (1) 0 ( ) ( ) exp( ) ( )A F C p p K C HdT F E hAT T T T dt V C RT VC\u03c1 \u03c1 \u2212\u0394 = \u2212 + \u2212 \u2212 \u2212 (2) equation (1) is the material balance equation, and equation (2) is the thermal balance equation, they represent the original dynamic model of the system, which is a typical set of nonlinear differential equations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002986_s00170-016-9507-2-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002986_s00170-016-9507-2-Figure11-1.png",
        "caption": "Fig. 11 FE model of the bed and base assembly",
        "texts": [
            " The impact hammer excites the assembly at the fixed driving point in the Z-direction. Twenty-three axis accelerometers are used to acquire the FRF data. The SIMOmethod is performed in the experiment to obtain all the FRFs. (iii) The acquisition bandwidth and frequency resolution are 520 and 0.5 Hz, respectively. The modal parameters of the assembly are obtained using the LMS Test.lab vibration analysis system. The first fiveorder modal results (modal frequencies and mode shapes) are shown in Table 5. The FE model of the assembly is shown in Fig. 11. The bed and base (structural components) are modeled by 28,233 hexahedral elements and 3176 pentahedral elements in the FE software-MSC.patran, and the material parameters of the components are shown in Table 3. The bolted joint (machine elements) between the bed and base is modeled by nine bolted joint elements. The stiffness matrices of the bolted joint elements are obtained using the same method as in Section 4.1 and are inserted into the finite element model through DMIG in MSC.patran. Executing normal model analysis, the modal results of the assembly, such as the natural frequencies and corresponding mode shapes, are obtained"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001090_pime_proc_1958_172_053_02-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001090_pime_proc_1958_172_053_02-Figure3-1.png",
        "caption": "Fig. 3. Two Flanks of a Screw Thread, Axis of Cylinder and Normals to the Flanks at Two Points of Contact --- Curves of intersection.",
        "texts": [
            " If, however, the cylinder contacts each flank at one point only then the rake correction for a ball and an equivalent cylinder can be shown to be identical. A formal analytical proof of this will not be attempted here but the equivalence of a ball and a cylinder at minimum distance from the thread axis for the case of a parallel external screw thread with equal flank angles will be demonstrated. Consider a cylinder applied to an external screw thread at any distance from the axis of the screw and at such an angle that it touches both flanks of the thread. Fig. 3 shows the two Aanks of the thread, the axis of the cylinder which is in a plane parallel to the plane of the figure and the normals to the flanks at the points of contact A and B. It should be noted that these normals are perpendicular to the axis of the cylinder and their projections in the plane of the figure are therefore parallel to one another and perpendicular to the straight line generators of the cylinder passing through A and B respectively. These generators are, in view of the postulated single contact, tangents to the flanks and, in particular, are tangents to the curves defined by the intersection of the flanks with the plane parallel to the plane of the figure and containing A and B. These curves of intersection are shown by a broken line in Fig. 3. If the cylinder is now pushed vertically downward into the thread and allowed to rotate so that it remains in contact with both flanks then the normals AA' and BB', which always remain in right sections of the cylinder, will approach one another. Eventually A' and B' will coincide and A and B will lie in a common right section of the cylinder. Fig. 4 shows the cylinder in this position. The right section of the cylinder in which A and B lie is identical to a section of a ball of equal diameter which would also contact the flanks at A and B"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002425_s11771-014-2021-5-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002425_s11771-014-2021-5-Figure7-1.png",
        "caption": "Fig. 7 Conjugate motion of slotting manufacture",
        "texts": [
            " 6, after active adjustment for the radial feed \u0394x2i and tangential feed \u0394y2i during machining, the position vector of the cosine gear and the gear slotting cutter can be coincide with each other. According to Eqs. (16) and (10), when the proposed condition is satisfied, the following relationship exists: 1 1 1 1 1 1 1 1 cos sin sin cos 1 i i i i i i i i x y x y                2 2 2 2 2 2 2 2 2 2 cos sin sin cos 1 i i i i i i i i x y x x y y a                   (20) J. Cent. South Univ. (2014) 21: 933\u2212941 938 3.3.2 Conjugate motion of slotting manufacture As shown in Fig. 7, P is the cutting point. If the cosine gear rotates an angle \u03c61i clockwise, the gear slotting cutter rotates an angle \u03c62i counterclockwise and moves \u0394x2i and \u0394y2i along x2 and y2, respectively. Moreover, every parameter is a function of the time dt. According to the conjugate principle [19], in order to ensure that the cosine tooth profile is tangent to the gear slotting cutter, the following equation can be satisfied at an arbitrary moment. 12 1 12 1( , ) 0o o o o p pf V n V n   (21) where 12 o mV denotes the relative velocity of the cutting point P in the coordinate system  (O, x, y)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002064_2016-01-1136-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002064_2016-01-1136-Figure1-1.png",
        "caption": "Figure 1. Cutaway of the e-LSD structure.",
        "texts": [],
        "surrounding_texts": [
            "The structure can be divided in three main parts: the right cartridge, the left cartridge and the housing. The latter, which is the central element, receives the engine torque by the outer crown fixed to this, and transfers it to the two solar gears; this is possible thanks to four planet bevel gears, which are connected to the central housing through the differential cross. The two bevel wheels, each one fixed to a half shaft and belonging to a different cartridge, generate an axial load due to their geometry; hence, in addition to radial bearings, each cartridge is equipped with an axial bearing to sustain this thrust. The structure of the right cartridge is more complicate as it contains the actuation system and the internal clutch. The actuation is composed by a main piston, placed in a suitable chamber, which is pushed by the oil in pressure; in detail, an external actuator increases the pressure using another piston activated by a ball screw mechanism and a servomotor. The main piston is constituted by two parts: one in contact with the friction discs and the other with the oil chamber. Thus, in order to allow a relative movement, an angular contact bearing is necessary. The internal clutch is formed by a discs\u2019 pack (oil immersed). All these discs are alternatively coupled with the differential cage and the half shaft through a splined coupling: this means that they are constrained for the revolution but free for the axial movements. The goal of the clutch is the generation of a friction torque that contrasts the relative rotation of the two half shafts: more torque is always delivered towards the slowest wheel, with the purpose to reduce the speed difference and to avoid a wheel spin. In addition to this, an unbalanced distribution of force on the same axis inevitably generates a yaw torque on the vehicle that modifies its attitude: the idea is to operate properly on the torque distribution to control the yaw angle of the vehicle, as well as the traction level [4]."
        ]
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.2-1.png",
        "caption": "FIGURE 6.2",
        "texts": [
            " For the vehicle dynamics task the mass, centre of mass position and mass moments of inertia of the vehicle body require definition within the multibody data set describing the full vehicle. It is important to note that the body mass data may include not only the structural mass of the body-in-white but also the mass of the engine, exhaust system, fuel tank, vehicle interior, driver, passengers and any other payload. A modern CAD system, or the preprocessing capability for example in ADAMS/View, can combine all these components to provide the analyst with a single lumped mass notwithstanding the cautions raised in Section 3.2.4. Figure 6.2 shows a detailed representation of a full vehicle model. In a model such as this there are a number of methods that might be used to represent the individual components. Using a model that most closely resembles the actual vehicle, components such as the engine might for example be elastically mounted on the vehicle body using bush elements to represent the engine mounts. The penalty for this approach will be the addition of 6 degrees of freedom (DOF) for each mass treated in this way. Alternatively a fix joint may be used to rigidly attach the mass to the vehicle body"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000174_978-0-387-92897-5_375-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000174_978-0-387-92897-5_375-Figure8-1.png",
        "caption": "Figure 8 shows an example of the calculated friction coefficient at the roller\u2013rib contact with ma = 0.13, moil = 0.02, and different values of asperity slopes defined in the x and y direction.",
        "texts": [
            " Annular orifices are sometimes employed for hole-entry hybrid hydrostatic/ hydrodynamic bearings or for shallow recess bearings. The annular orifice gives bearing film stiffness that is two-thirds of the bearing film stiffness of a simple orifice bearing. When combined with a shallow-recess bearing, the combined stiffness is increased. The discharge coefficient of an orifice depends on the Reynolds number; see typical values in Fig. 6 for an orifice with l/do less than 1. Supply Ps Diaphragm Gap z pr Diaphragm valve pr Journal bearing Fig. 8 Rowe valve for journals and opposed pads Constant flow Constant flow control is attractive for a single-pad bearing. It only requires a constant flow pump or a constant flow valve in the supply line. However, for a large number of recesses, the cost of several pumps or valves makes this a costly approach. Although the concept of a supply pressure may seem strange for constant flow, it is a real issue. A maximum supply pressure must be ensured at which constant flow can be maintained. The maximum supply pressure governs the maximum load, as for other types of flow control",
            " Pressure-sensing valves Pressure-sensing valves optimized for maximum stiffness increase flow rate in proportion to bearing load. This maintains clearance constant within the limitations of system linearity and yields virtually infinite static stiffness of the bearing film. There are a number of such designs based on spool valves and diaphragm valves (O\u2019Donoghue and Rowe 1969a, b). The Mohsin valve was developed for single-pad bearings (Mohsin 1962), Fig. 7, and the Rowe valve for journals and opposed pads, Fig. 8 (Rowe and O\u2019Donoghue 1968). Single diaphragm valve, Fig. 7 Bearing recess pressure acts on the diaphragm, causing it to deflect. The deflection allowsmore flow into the heavily loaded recess. By suitable selection of diaphragm stiffness or by adjustment of supply pressure, it is possible to adjust the bearing film stiffness. The flow restriction offered by a diaphragm valve is as given above for the annular restrictor where the diaphragm valve film thickness is z. If the diaphragm is very stiff, the valvemakes a laminar annular flow restrictor having the same properties as a capillary",
            " The diaphragm deflection is linear with load since the deflections are very small compared with the diaphragm diameter. The diaphragm film thickness is therefore given by z \u00bc zo \u00feWd=ld , where the deflection force Wd \u00bc Ad\u00f0pr b\u00de for a single-sided diaphragm valve and zo is the diaphragm film thickness when the bearing film thickness h=ho, the design condition. The dimensionless diaphragm stiffness is defined as ld \u00bc d pr d z , where z \u00bc z=zo and pr \u00bc pr=Ps . Infinite film stiffness of singlepad bearings is obtained when ld \u00bc 3b\u00f01 b\u00de, leading to ld = 0.75 when b \u00bc 0:5. Double diaphragm valve, Fig. 8 A double diaphragm restrictor controls two recesses with one diaphragm (Rowe 1966, 1969). The design is easy to make, and one restrictor block using a single diaphragm can be designed to supply a number of recess pairs. For opposed-pad bearings and journal bearings, the diaphragm must be made twice as stiff as for a single-pad bearing to achieve a similar optimum condition. This is because a change in pressure occurs on both sides of the diaphragm. The deflection force on the diaphragm is Wd \u00bc Ad\u00f0p1 p2\u00de, where p1 and p2 are the opposing recess pressures",
            " This is illustrated in a plot of shear stress versus shear rate (Fig. 7) where oils A and B have about the same viscosity, but B has a measurable yield stress. Full scale engine studies in the 1960s and 1970s showed two distinct pumpability failure modes in engines at low temperatures (ASTM DS-57 1975; Stewart 1968; Shaub 2000). In some cases, the oil viscosity was too high to allow oil to flow to the pump. Visual observation of the oil pan showed a depression in the oil level over the pump inlet, and oil pressurization showed a slow steady increase (see Fig. 8a, b). On the inlet side of the oil pump, sufficient force was generated to produce a significant vacuum in this tube. For this flow-limited behavior, the critical path was shown to be along the inlet tube ahead of the pump, and FLOW-LIMITED PUMPABILITY CONDITION Critical Flow Path: Pump Inlet T a Rheological Measurement Methods and Equipment, Fig. 8 (a, b as detected by oil pressure in the oil circuit (b) Shear Rate S h ea r S tr es s B A Yield Stress Oil B Shear stress-shear rate plots for oils with and without yield stress correlations to viscometer measured behavior were best with a defined shear stress of 525 Pa. With other engine/ oil/temperature combinations, a second pumpability failure mechanismwas identified inwhich a channel formed in the oil when oil in the pan could not slump fast enough to replace the initial amount drawn to the pump (illustrated in Fig",
            " 7, the punch surface contacts incompletely to the matrix. At this punch end E, the stress components sr \u00bc try \u00bc 0, therefore, the end is a smooth contact and not a singular point. The P mP a Rigid Punch Problem with a Crack, Fig. 6 Circular-ended inclined punch y x P mP E a\u2032 Rigid Punch Problem with a Crack, Fig. 7 Flat-ended inclined punch with incomplete contact restraining condition for the smooth contact at the end E (s \u00bc b) is i\u00f01 im\u00de\u00f01 a\u00de 2pw\u00f01\u00de\u00f0b a\u00de P \u00fe \u00f01 m\u00def \u00f0b\u00de \u00bc 0 \u00f015\u00de Rigid Punch Problem with a Crack, Fig. 8 Wedge-ended punch with incomplete contact P mP a\u2032 Rigid Punch Problem with a Crack, Fig. 9 Circular-ended inclined punch with incomplete contact P mP Rigid Punch Problem with a Crack, Fig. 10 Circular-ended vertical punch with incomplete contact R The function f \u00f0s\u00de in (15) is given by f \u00f0s\u00de \u00bc G R \u00f01 im\u00de k\u00fe 1 E2 0 m 1 a \u00fe 1 m 1 b \u00fe 2E0Ec 1 w\u00f01\u00de\u00f01 s\u00de \u00fe E2 0 w\u00f01\u00de\u00f01 s\u00de2 \u00fe XN k\u00bc1 2E0Ek zk 1 1 w\u00f01\u00de\u00f01 s\u00de \u00fe 1 w\u00f0zk\u00de\u00f0zk s\u00de \u00fe XN k 6\u00bc1 EkE1 zk z1 1 w\u00f0z1\u00de\u00f0z1 s\u00de 1 w\u00f0zk\u00de\u00f0zk s\u00de \u00fe XN k\u00bcl E2 k w\u00f0zk\u00de\u00f0zk s\u00de2 \u00fe XN k\u00bcl E2 k m zk a \u00fe 1 m zk b \u00fe 2EcEk 1 w\u00f0zk\u00de\u00f0zk s\u00de 2 6666666666666664 3 7777777777777775 \u00f016\u00de The contact length a0 is obtained, a0 \u00bc o\u00f0a\u00de o\u00f0b\u00de \u00f017\u00de 1. Flat-ended inclined punch with incomplete contact (Fig. 7) The stress function is presented by (2) with H3\u00f0z\u00de \u00bc 0\u00f0R \u00bc 1; e 6\u00bc 0\u00de. The vertical force is applied on the y axis (e = 0). Also the coordinate b of the smooth end E must be determined, satisfying (15) by iterative calculation. The contact length is calculated by (17). (Hasebe et al. 1999) 2. Wedge-ended punch with incomplete contact (Fig. 8) The solution is the same as that of (2). The wedge angle e is known. The vertical force is applied on the y axis (e = 0). 3. Circular-ended inclined punch with incomplete con- tact (Fig. 9) P mP The stress function is presented by (2) \u00f0R 6\u00bc 1; e 6\u00bc 0\u00de. The vertical force is applied on the y axis (e = 0). The inclined angle e is determined by Rm \u00bc 0 in (14). Also the coordinate b of the smooth end E must be determined, satisfying Eq. (15) by the iterative calculation. The contact length is calculated by (17) (Hasebe and Qian 1997, 1998)",
            " The mass rate plot is perhaps the most indicative of the damage severity; however, after inspection and maintenance several particles will remain trapped within the assembly. As a result, when the repaired assembly is initially started a rush of debris will pass the sensor, even though the newly repaired bearing may not be damaged. The mass rate plot also helps the user determine whether damage is accumulating or has been arrested. 100 Sample A Sample B Composition Si P Sample C Sample D Contaminants in Various Oil Samples 200 300 C on ce nt ra tio n Le ve l ( m g/ m l) 400 500 0 S Ca Fe Zn Rolling Bearing Condition Monitoring, Fig. 8 Lubricant particle composition from an automotive application (Ai 2001) R Some limitations of debris monitoring are as follows: \u25cf Oil debris monitors are only able to detect particles that pass through the sensor. Any particles that adhere to the side of the housing or gear box are not detected. \u25cf Oil debris monitors have the ability to detect ferrous particles and nonferrous particles; however, the particlesmust have the ability to respond to amagnetic field. \u25cf If several components are manufactured from the same material it may be difficult to determine which component is damaged, but it is a good indicator that maintenance is required",
            " Analytical ferrography samples suspended particles from lubricant by placing a magnet beneath a test specimen slide. The magnet and slide are askew such that larger particles collect first on the weak magnetic side and smaller particles collect on the strong magnetic slide. A trained technician is then able to determine the quantity, size, composition, and morphology of the ferrous and nonferrous wear particles. 5. X-ray techniques yield the elemental composition of the debris. This allows a user who has sampled oil from the system to determine the type of contaminate in the oil (Fig. 8). The debris type is then compared with the know system materials such as the gear or bearing steel. For further information, see (Humphrey 1996; Humphrey et al. 2002a, b). There numerous condition monitoring techniques. Many articles have been written comparing one device with another and many more discuss data analysis. Industrial monitoring is advantageous and the devices mentioned here can be implemented. Temperature monitoring is the most frequently used method, mainly due to ease of implementation and interpretation",
            " It is calculated on the most loaded roller\u2013 cone contact: L \u00bc hP Ra : L 2:b 0:56 with L\u00bc 0:03 inch or 0:762 mm \u00f013\u00de Figure 6 shows the lube factor calculated versus the modified L ratio. The profile and misalignment factor accounts for edge stresses that can be found at the edge of a roller\u2013race contact, especially if misalignment between inner and outer race (and hence, roller and race) occurs. Note that complex multi radii shape can be defined for reducing edge stresses. Figure 7 shows an example of calculated profile and misalignment factor. The low load factor can be defined as a function of P/C90 lubrication and profile and misalignment factors, as shown in Fig. 8. Finally, debris factors can also be calculated as a function of the applied bearing load, indentation slope, and density, see Fig. 9 from Ai (2001). There are no specific factors describing residual stress and hoop stress effect on life, but these effects can be easily studied using stress-based models. The heat treatment process associated with case carburized steel (as opposed to through hardened steel) creates compressive stresses (negative stresses on the order of \u2013200 MPa) that contribute to increase the magnitude of the negative hydrostatic pressure, hence increasing the endurance limit tu described by (11) and final life",
            " Specimens fail under maximum Hertzian contact stress that is less than 4.41 GPa. Avibration sensor is used to end the test when a sharp increase of vibration level is detected. This is a standard test rig to produce rolling contact fatigue. This test rig has been used to observe the peeling mode of failure and surface-originated flaking process of the bearings (Tokuda et al. 1982). Two specimens with different dimensions are used to generate line contact. Their diameters are 53 and 50 mm for upper and lower specimen, respectively. Figure 8 depicts the schematic construction of the essential parts of the rig. Slide-to-roll ratio at the contacts was controlledusingdifferent diameter rings.Also the contact ellipses with different aspect ratios were controlled using different curvature radius rings. The lower specimen uses a constant rotational speed, 2,000 RPM, while the upper specimen uses variable rotational speeds. Variable applied loads are used to generatepeelingandflaking.ThemaximumHertziancontact stress ranges from 0.98 to 3.92 GPa. ing roll 6204 6204 pecimen 2 i22 ball e roll ion of the essential part of the rig (Ito et al. 1982) R Oil supply Upper roll 6306 Supporting roll Lower roll N308 NU306 Driving roll Driv e Specimen f12xi12 f20xi20 Cylinder Rolling Bearing Test, Fig. 7 Schematic construction of the essential part of the test rig (Sugiura et al. 1982) Center shaft Specimen Specimen Load Spherical bearing Rolling Bearing Test, Fig. 8 Schematic construction of the essential part of the rig (Tokuda et al. 1982) LOAD DUE TO SPRING LOADING ASSEMBLY FLUID DRIP OR FLOW FEED 4 - THERMO COUPLE HOLES DRIVEN END BRG.1 BRG.2 OUTLET BRG.3 BRG.4 Rolling Bearing Test, Fig. 9 A section of the four-bearing fatigue life test rig (Hobbs 1982) SPLASH GUARD TEST LUBRICANT BALL RACE THRUST BEARING LOAD SPACER CHUCK (FITS INTO ROTATING SPINDLE) Rolling Bearing Test, Fig. 10 Rolling four-ball test layout (Eastaugh 1982) R A test rig contains four size 6,208 bearings",
            ", MC is maximum and MP is nil. Under a combined axial and radial load, the contact angle gets larger and E gets rapidly larger than 1. One rolling line is then found (at y 0) and the factor fc is nil while fp is equal to 1, i.e., MP is maximum and MC is nil. In TRB, sliding speed is almost nil at the roller\u2013race contact, but slip is found at the roller\u2013rib contact where A=ratio.LApex Z LApex Y A f f 2b Frib Mrib X Z Rolling Bearing Torque, Fig. 7 Friction and moment at roller \u2013 rib contact Rolling Bearing Torque, Fig. 8 Roller \u2013 rib friction coefficient, function of L and asperity slope R a small load Qrib is applied. As a result, a sliding force Frib and braking rib moment Mrib (both shown in Fig. 7) can be calculated. The rib friction force Frib and moment Mrib can be calculated approximately using: Frib mrib:Qrib Mrib mrib:Qrib:f : ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1\u00fe 0:18 b f 2 s \u00f012\u00de where mrib is the local friction coefficient and Qrib the small roller\u2013rib load calculated approximately as: Qrib 2:Q: sin n \u00f013\u00de where n is the half included roller angle, on the order of 2 ",
            " f rail b d f rail Railbreak Growing cracks a rail in the longitudinal direction showing development of RCF g b palling and (b) rail break R Rolling Contact Fatigue (RCF), Fig. 4 Longitudinal cross section of a rail with head checks (Railtrack 2001) Rolling Contact Fatigue (RCF), Fig. 5 Schematic of gauge corner cracking Rolling Contact Fatigue (RCF), Fig. 6 Extensive fatigue cracking above gauge corner (Garnham et al. 2007) 10 mm 20 mm a b Rolling Contact Fatigue (RCF), Fig. 7 (a) Multiple squats on rail, (b) severe squat (Li et al. 2008) 10 mm Rolling Contact Fatigue (RCF), Fig. 8 Appearance of tongue lipping New Wheel Hollow TreadWear False Flange New wheel Hallow wheel a b Rolling Contact Fatigue (RCF), Fig. 9 False flange development on a wheel by wear Head checks are surface defects and form like small hairline cracks on the railhead, as seen in Fig. 4. They are usually found at the wheel/rail contact locations. On straight track and in curves with gentle curvature, HC occurs at the top of the railhead, and for sharper curves, HC develops at the gauge corner. In gauge corner cracking, cracks develop at the corner of the rail",
            " The appearance of GCC on the rail is shown in Figs. 5 and 6. Rolling Contact Fatigue (RCF), Fig. 10 False flange damage on rail Rolling Contact Fatigue (RCF), Fig. 11 Typical wheelburn R Rail squats are cracks that grow horizontally below the running surface and are thought to be caused by heavy wheel load over a cracked rail. This kind of RCF defect looks like a depression mark on the railhead (Fig. 7). The stresses due to trains passing over a rail deform the material and extrude it into a tongue, as shown in Fig. 8. During this process cracks initiate and develop at the interface between extruded material and non-extruded material. A new wheel has a conical tread and a flange on the inner side, as shown in Fig. 9a. After some period of operation, the wheel wears and a hollow may develop on the tread. Toomuch hollowing (because of excessive wear) can cause a false flange to develop on the outer side of the wheel (Fig. 9b). This damages the railhead, as seen in Fig. 10. Wheelburns result from frictional heating produced by a wheel spinning at the same spot on the rail head",
            "00 V er tic al C on ta ct F or ce ( N ) Rolling Contact Fatigue in Rail \u2013 Insulated Rail Joints (IRJ), Fig plan (top) view) Two strain-gauged IRJs were installed in an Australian heavy haul network; some data are presented from one of the strain-gauged test specimens, where the advantage of the localized nature of the contact zone characterized by plasticity is exploited. As large surfaces in both the wheel and the rail remain under low elastic level stresses, ample opportunity for strain gauging some key locations (Fig. 8) existed. The IRJ test specimen was factory fabricated with the usual thermal treatment. Strain gauging was performed in coordination with the manufacturer to ensure safe and reliable location of the gauges in the key spot identified by the FE analysis. Only the rail was strain gauged. Straingauge rosettes were used on both sides of the rail web on both rails that form the IRJ. The two rails were also strain gauged at the foot for the longitudinal strains. A total of fourteen strain channels were thus recorded",
            "45 ms to cross a nominally 10 mm thick joint), data were recorded at a very high frequency of 20,000 data points per second. To minimize the volume of data, an ultrasonic object detector was used to trigger the data recorder. When no trains were present, the recorder recorded the strains at the rate of one data point per 5 min. The time and temperature (ambient) readings were also recorded at this rate. 200 300 400 ce (mm) 174KN N Endpost of rail head .006 1.008 1.010 1.012 me (S) . 7 Rail/wheel contact force history (including railhead mesh\u2013 R Fig. 8 Rail showing highly strained locations Fig. 9 Field experimentation of the IRJ Field data FE model 600 500 400 300 200 100 0 0 2.5 5 7.5 1.0 2.5 Time (Millisecond) V er tic al s tr ai n (M ic ro st ra in ) Fig. 11 Vertical strain signature predicted FE and filed test: loaded wagons The corridor services mixed traffic. As the project was focused on the response of IRJ to heavy axle loaded traffic, strategies were required to sort out the data corresponding to the loaded trains. Several strategies were formulated based on the information on train composition, traffic data, and the strain signatures",
            " l and shear deformation of lip, H, q and g) lculate , load and power and output No No kewed flow factors and conductivity w factors and skewed flow factors distribution by solving Reynolds eqn.) for P, H, d, g and Y operating parameters sure (calculated from a finite element model) tical approach (Rocke and Salant 2005) R 0.0035 0.0040 0.00450.0030 0.000 \u20130.002 0.002 \u20130.004 0.004 \u20130.006 0.006 \u20130.008 0.0025 Q * V* 5 x-oriented asperities 5 circular asperities 5 y-oriented asperities 0.0020 Rotary Lip Seal Analysis, Fig. 8 Dimensionless reverse pumping rate versus dimensionless speed (Salant and Flaherty 1995) 0.0 0.88 0.66 0.00 0.0 0.2 0.4 1.0 0.8 X** Y* 0.6 0.22 0.44 0.3 0.6 0.9 1.2 P** Rotary Lip Seal Analysis, Fig. 9 Dimensionless fluid pressure distribution (Salant and Flaherty 1995) The two-dimensional energy equation, with appropriate boundary conditions, is solved for the temperature field (Day and Salant 1999). In the second approach, the film is treated as a line heat source, and an average film temperature is computed using a simple thermal balance",
            " They can be broadly classified as: \u2013 Seal only \u2013 Radial dust lip \u2013 Multiple radial dust lips \u2013 Axial dust lips \u2013 Multiple radial and axial lips \u2013 Labyrinth and flinger features with any of the above The simplest form of excluder to use is a radial dust lip, molded into the heel of the seal (Fig. 7). These are intended to protect the main seal lip from a light dust environment. The favored approach with such a wiper is to ensure that it is lubricated and consider a small relief on the lip to allow venting to prevent a vacuum formation. + Rotary Shaft Lip Seals, Fig. 7 A simple excluder, a lip seal with a single radial dust lip + Rotary Shaft Lip Seals, Fig. 8 Lip seal and axial excluder in a cassette arrangement that will permit high speed operation with positive exclusion at low speed R A felt dust lip can act as a filter and remove dust particles but will allow the space between the seal and dust lip to breath, hence removing the problem of creating a vacuum between the seal and dust lip. Felt wipers can be effective in dry, dusty conditions but are not satisfactory for excluding water from the seal. Axial dust lips can be very effective as part of the exclusion system",
            " The shaft flange will provide a centrifugal effect to any contaminant in the seal area and will disperse much of any contaminant present when the shaft is rotating. There is also the tendency to disperse any wear debris created from the dust lip, which further helps to keep the seal area clear. Designs with a rotating flexible sealing element will subject it to centrifugal force, which will tend to lift the seal from contact during rotation. The excluder lip can therefore be designed to have a higher contact stress when static, to prevent ingress of contaminant, but will operate primarily as a centrifugal seal when the shaft is rotating at high speed (Fig. 8). A cassettes is an integrated seal and shaft sleeve assembly and may also include many of the exclusion features (Fig. 8). The benefits can include: \u25cf The seal supplier is providingboth the static anddynamic components of the seal in one controlled assembly. \u25cf The unit is pre-assembled in controlled conditions so there is less opportunity for damage to the seal lip during transport, storage, and assembly. \u25cf The seal supplier provides the shaft surface texture and hardness required. \u25cf A complex seal geometry with optimized excluder design incorporating radial flanges and axial sealing lips can be effectively assembled",
            " A set of solutions for contact ellipticity of k = 2 with the transverse roughness are given in Fig. 7 as examples, demonstrating the entire transition k= k=1.0 k=1/2 the contact ellipticity, the less the lateral flows, and the stronger with varying rolling speeds and resulting l ratios. Solutions for other contact types and roughness orientations are similar in nature. In the following, the roughness orientation effect on lubrication is discussed for different types of contact geometry and results are summarized in Fig. 8. First, line contact cases are analyzed using a three-dimensional mixed EHL model recently developed by Ren et al. (2009) in order to take into account the effect of threedimensional roughness commonly found in engineering applications (also refer to \u201c\u25b6 3D Line Contact EHL\u201d). It can be seen from the upper-left graph in Fig. 8 that in line-contact EHL the transverse roughness yields thicker EHL film than the isotropic and longitudinal under the same operating conditions. This orientation effect is significant in the mixed EHL region, where both lubricant films and asperity contacts coexist and neither can be ignored (the contact load ratio is roughly in a range of 0.1 < Wc < 0.9, as illustrated in Fig. 9). It is relatively insignificant at extremely low speeds due to vanishing hydrodynamic action, and also at very high speeds, which lead to vanishing asperity contacts and strong hydrodynamics with the lubricant film thickness significantly greater than the composite roughness",
            " When the speed approaches zero, L = hcs/s approaches zero in the meantime, but the decrease of l = ha /s is much slower due to the existence of rough asperities that are difficult to be completely flattened. That is why in Fig. 10 the ratio of ha/hcs may increase as the L ratio goes below 0.5. The situation for point contact cases may become more complicated. When the contact ellipticity, k = b/a, is large, the situation is close to that of line contact due to relatively insignificant lateral flows in the EHL conjunction. The basic trend for the surface roughness orientation effect may be similar to those of line contact, as shown in Fig. 8 for the cases of k = 2.0. If the contact ellipticity is small, e.g., k = 0.5, the lateral flows become significant and the entraining action is much weakened. Since the transverse roughness enhances the lateral flows and the longitudinal enhances the entraining action in the surface motion direction, the longitudinal roughness becomes more favorable, yielding thicker lubricant films, as indicated in Fig. 8 for the cases of k = 0.5. In the circular contact cases, the lateral flows and the entraining flows are quantitatively comparable. The roughness orientation effect on film thickness, therefore, appears to be less significant under the same operating conditions, but the longitudinal roughness is still relatively more favorable for lubrication formation in the analyzed cases. It is of extreme importance inmachine element design and production how to select, optimize, and specify surface finish. Today, requirements for continuously improving Line Contact Isotropic Transverse Longitudinal 10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000984_j.optlastec.2014.11.015-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000984_j.optlastec.2014.11.015-Figure3-1.png",
        "caption": "Fig. 3. Cross-section of the joined components: welding configuration varies from the reference position (beam axis orthogonal to the surface, pointing the intersection of AISI 430F and AISI 440C) with the use of input parameters.",
        "texts": [
            " It is worth noting that, even if some micro-cracks are present, samples belong to the class 1 according to the ISO 13919-1:1996 standard and thus are accepted for the purpose [16]. To avoid any presence of cracks, which mostly develop on the side of the martensitic steel, it is here hypothesized to change the position of the laser beam towards the ferritic steel, by offsetting the beam axis in respect to the interface of the two materials. This is done to compensate the difference in supplied energies between the two materials showing dissimilar absorption of the laser beam and different thermal conductivities. Nevertheless, increasing the beam offset (y in Fig. 3) may result in resistance length under the design value with a consequent reduction in the mechanical strength. To avoid this limitation and, at the same time, to obtain a weld profile less prone to surface-cracking, the beam incidence angle \u03b8i (see Fig. 3) has to be varied as well. This results in the need of defining weld cross sectional geometries for each combination (y, \u03b8i) adopted. Inclining the laser beam has a further beneficial effect of increasing the irradiated area which becomes elliptical (increasing \u03b8i) with the longer axis disposed along y-direction. According to previous experiences reported in [17] this elongation of the irradiated area, allows for a reduction of the thermal gradient in y-direction which may contribute to a less severe thermal cycle on the extinction zones of the weld bead, where micro-cracks usually appear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003196_cistem.2014.7076970-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003196_cistem.2014.7076970-Figure9-1.png",
        "caption": "Fig. 9 .Structure du moteur avec la culasse statorique",
        "texts": [
            " Cette variation d'induction \u00e0 Z = 0 est comprise entre - Bmax =-1.9T et + Bmax = 1.9T, ceci est visible sur la Fig. 8-a. Selon les r\u00e9sultats, on conclut que, pour le m\u00eame ensemble de param\u00e8tres et les m\u00eames conditions, la valeur moyenne de l'induction magn\u00e9tique radiale de l'inducteur avec du fer est meilleure que celle de l'inducteur sans fer, et donc on adopte la structure de l'inducteur avec un mat\u00e9riau ferromagn\u00e9tique dans l'\u00e9tude suivante. Nous ajoutons la culasse du stator \u00e0 l'inducteur comme on le voit sur la Fig. 9-A. Comme il est montr\u00e9 dans la Fig. 7 la courbe de l'induction radiale (Br) de l'inducteur diminue \u00e0 mesure que nous nous approchons du sol\u00e9no\u00efde et on remarque que l'induction devient n\u00e9gative pour Z>6cm. Cette inversion du champ le long de la longueur axiale de la machine n'est pas favorable pour la force \u00e9lectromotrice de l'induit. Pour cela, il est propos\u00e9 de r\u00e9duire la longueur axiale \u00e0 chaque extr\u00e9mit\u00e9 du stator comme cela est pr\u00e9sent\u00e9 sur la Fig. 9-B. La simulation a \u00e9t\u00e9 r\u00e9p\u00e9t\u00e9e pour d\u00e9terminer quelle longueur de culasse statorique nous donne l'induction radiale maximale. Puis la structures r\u00e9duite \u00e0 \u00e9t\u00e9 compar\u00e9e \u00e0 la structure initiale de r\u00e9f\u00e9rence ou l'induit occupe toute la longueur entre les deux sol\u00e9no\u00efdes. La Fig. 10 repr\u00e9sente deux courbes de l'induction magn\u00e9tique radiale Br (Z). Apr\u00e8s avoir calcul\u00e9 la valeur moyenne de l'induction radiale utile, pour chaque configuration, on obtient une induction de 1,72 T dans la topologie A et 2,06 T pour B (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000784_ilt-01-2012-0013-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000784_ilt-01-2012-0013-Figure1-1.png",
        "caption": "Figure 1 Squeeze-film geometry between a sphere and a porous plane surface with roughness in the presence of a transverse magnetic field",
        "texts": [
            " They found that in the MHD squeeze-film pressure, load-carrying capacity increases in the presence of externally applied magnetic field, and for large values of Hartmann number, the response time increases. In the current study, we analyse the effects of surface roughness on the MHD squeeze-film characteristics between a sphere and a porous plane surface which have not been studied so far. Hence, in this paper, an attempt has been made to analyse the combined effects of surface roughness and permeability on the MHD squeeze-film characteristics between a sphere and a plane surface. The physical configuration of a rough porous squeeze-film bearing is shown in Figure 1. A solid sphere of radius R is approaching the rough porous bearing of wall thickness H0 with a velocity h/ t under a constant load. An incompressible electrically conducting fluid is considered as a lubricant in the film and also in the porous region between two surfaces. An externally uniform transverse magnetic field B0 is applied to the lower surface. The stochastic film thickness H is represented by: H h hs(r, , ) where h hm r2/2R provided r R denotes the nominal smooth part of the film geometry, while hs is the part due to the surface asperities measured from the nominal level and is a randomly varying quantity of 0 mean and is in index describing the definite roughness arrangement; hence, for a given value of , the surface roughness component hs of the film thickness becomes a deterministic function of the space variables"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001811_0954405415608784-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001811_0954405415608784-Figure1-1.png",
        "caption": "Figure 1. Schematic of test rig for the ball bearings.",
        "texts": [
            " The influence of speed on the internal load distribution and ball bearing contact angles was also analysed. After obtaining the internal load, the optimum preload could be determined by changing the reliability factor within the fatigue life model. The proposed method to determine optimum preload was verified using measured performance indicators, such as the temperature, motor currents, and vibration of ball bearings. The results showed that the optimum preload contributed to ideal behaviour in the test ball bearings. Figure 1 illustrates a schematic of the experimental system that was used to verify the optimum preload for ball bearings. The test shaft is driven by a motorised spindle and is supported by ball bearings (B7007C), which are acted on by hydraulic parts. The hydraulic pressure induces displacement in the axial direction, pushes the outer ring of the rear bearing, and eventually transfers this displacement to the inner ring of the rear bearings through the shaft. Therefore, the displacement caused by hydraulic pressure is converted into axial force"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure15-1.png",
        "caption": "Fig. 15. Electric field distribution around the energized parts around the conductor car",
        "texts": [
            " In this case finite element solution of COMSOL MultiPhysics was used to determine both the electric potential and the electric field distribution in the vicinity of the arrangement. The first part of the simulation was the determination of the electric potential of the different elements of the arrangement. Fig. 13 shows the values of the electric potential during the most critical conditions while the sinusoidal waveform of the voltage reaches its peak value. The nominal (RMS) value of the system voltage was 400 kV. From the electric potential values electric field distribution can also been determined. Fig. 14 and Fig. 15 shows the electric field distribution of the arrangement and in the vicinity of the conductor car. 978-1-5090-5165-6/16/$31.00 \u00a92016 IEEE From the results it can be determined that during practical working conditions peak value of electric field in the air in the surroundings of the arrangement is about 55 kV/m. Calculated with a 20 kV/cm value as the dielectric strength of the homogenous electric field and considering the inhomogeneity factor as 10, safety factor regarding to the electrical stresses is above 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.8-1.png",
        "caption": "FIGURE 8.8",
        "texts": [
            "81 with respect to time t yields Fq q; t \u20acqh 2 6666664 1 0 0 0 0 0 0 1 0 0 0 0 0 0 0 0 1 0 1 0 \u20181 sin q1 1 0 \u20182 cos q2 0 1 \u20181 cos q1 0 1 \u20182 sin q2 0 0 1 0 0 0 3 7777775 2 6666664 \u20acx1 \u20acy1 \u20acq1 \u20acx2 \u20acy2 \u20acq2 3 7777775 \u00bc 2 66666664 0 0 0 _q 2 1\u20181 cos q1 \u00fe _q 2 2\u20182 sin q2 _q 2 1\u20181 sin q1 _q 2 2\u20182 cos q2 0 3 77777775 (8.83) \u20acq can be solved using Eq. 8.83. This is left as an exercise (Exercise 8.7). With the Cartesian generalized coordinates (or absolute coordinates)dthat is, position and orientation of every single body in the systemda large system of equations is obtained. For example, for a slider-crank mechanism that consists of three moving bodies, shown in Figure 8.8(a), the absolute coordinates are q \u00bc x1; y1; q1; x2; y2; q2; x3; y3; q3 T (8.84) However, using Cartesian generalized coordinates, the constraint equations can easily be formulated automatically, which is ideal for implementation on computers. There is a simpler set of generalized coordinates that can be chosen for problem formulation. For the system shown in Figure 8.8(b), the joint coordinates that correspond to kinematic joints in system can be chosen as q \u00bc \u00bdqA; qB; Z3 T (8.85) where qA and qB correspond to the rotation DOF of the two revolute joints A and B, respectively. Z3 represents the translation DOF of the slider joint at point C. Note that the parameters defined in Eq. 8.85 are consistent with those of Example 8.5. Assuming the rotation angle of the crank is prescribed by a driving constraint qA \u00bc f(t) \u00bc ut, the constraint equations of the system can be formulated in terms of q \u00bc \u00bdqA; qB; Z3 T as F\u00f0q; t\u00de \u00bc 2 64 \u20181 cos qA \u00fe \u20182 cos qB Z3 \u20181 sin qA \u00fe \u20182 sin qB qA ut 3 75 \u00bc 0 (8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002074_hnicem.2015.7393169-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002074_hnicem.2015.7393169-Figure2-1.png",
        "caption": "Fig. 2. An schematic diagram of a 3-arm cooperating parallel manipulator, with the corresponding reference frames and the relative position vectors.",
        "texts": [
            " The term \u201cparallel\u201d may have been a misnomer because this normally refers to manipulators with end-effectors that are rigidly connected to each together. However, we use the term \u201cparallel\u201d in this sense that the bases of the manipulators are rigidly connected to each other, as opposed to the \u201cseries\u201d connection where a manipulator base is rigidly connected to the end-effector of another manipulator. Other manipulator kinematics study include [16], [17]. The naming convention for most symbols used in this work are shown in Table I. Based on the schematic diagram of the 3-arm cooperating parallel manipulators in Fig. 2, the reference frames are assigned. The base reference frames are odd-numbered, while the end-effector reference frames are even-numbered. Relative position vectors connect the endeffectors. Consider reference frames {i} and { j}, such that i p j is the position of frame { j} with respect to frame {i}, and iR j is the rotation of frame { j} with respect to frame {i}. In addition, a Jacobian iJ j can be expressed with respect to those frames. From the figure, we state the following conventions for the Jacobians of the standalone manipulators",
            " We assign the position Jacobian iJp j and orientation Jacobian iJo j as components of the Jacobian iJ j, that is, iJ j =[ iJp j, iJo j ]T . The joint velocities q\u0307i j = [q\u0307i, q\u0307 j] T , such qi and q j are the joint velocities of the robot with end-effector frames {i} and { j}, respectively. For example 1J2 = [1Jp2, 1Jo2] T is the Jacobian for robot A, and 2J4 = [2Jp4, 2Jo4] T is the relative Jacobian of the dual-arm consisting of robots A and B. The dual-arm joint velocities q\u030724 = [q\u03072, q\u03074] T , where q\u03072 are the joint velocities of robot A q\u03074 are the joint velocities of robot B. Based on the frame assignment shown in Fig. 2, we present here the modular relative Jacobians for dual-arms as derived in [3]. The relative Jacobian for a dual-arm consisting of robots A and B is 2J4 = [ \u22122\u03a84 2\u21261 1J2 2\u21263 3J4 ] , (1) the relative Jacobian of a dual-arm consisting of robots B and C is 4J6 = [ \u22124\u03a86 4\u21263 3J4 4\u21265 5J6 ] , (2) and lastly, the relative Jacobian for dual-arm robots A and C is 2J6 = [ \u22122\u03a86 2\u21261 1J2 2\u21265 5J6 ] . (3) Such that the wrench transformation matrix i\u03a8 j is defined as i\u03a8 j = [ I \u2212S(i p j) 0 I ] (4) and the rotation matrix i\u2126 j is expressed as i\u2126 j = [iR j 0 0 iR j ] . (5) Given \u03c9 = [\u03c9x,\u03c9y,\u03c9z] T , the operator S(\u03c9) is the skew symmetric operator used to replace the cross-product operator and is expressed as S(\u03c9) =  0 \u2212\u03c9z \u03c9y \u03c9z 0 \u2212\u03c9x \u2212\u03c9y \u03c9x 0  . (6) To complete the definition of the modular dual-arm manipulators the shown robots in Fig. 2, we define the relative position vectors between the end-effectors, called i p j for the paired robots. We express them here as 2 p4 = 2R1( 1 p3 + 1R3 3 p4 \u2212 1 p2) 4 p6 = 4R3( 3 p5 + 3R5 5 p6 \u2212 3 p4) 2 p6 = 2R1( 1 p5 + 1R5 5 p6 \u2212 1 p2). (7) To derive the modular relative Jacobian for the 3-arm cooperating parallel manipulators, we invoke the approach used in [3], that is, we express translational and rotational velocities of the end-effectors with respect to each other. Thus the relative position of frame {6} with respect to frame {2} is expressed as 2 3 p6 = 2 p4 + 2R4 4 p6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003668_978-3-319-06698-1_19-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003668_978-3-319-06698-1_19-Figure1-1.png",
        "caption": "Fig. 1 a A serial kinematic chain of segments with marked connections between two neighboring segments and IMU\u2019s measured quantities. Quantities denoted with subscripts j refer to the j-th segment. Quantities marked with left superscripts I, S, and E are expressed in the coordinate frame of the IMU, segment, and Earth, respectively. b RL Right\u2013left and LR left\u2013right sequence of orientation calculations, denoted with numbers in circles, when the right or the left foot is in stance phase, respectively",
        "texts": [
            " The sensory fusion for estimation of segment orientations in long-term and dynamic motion proposed in this chapter is built upon an extended Kalman filter (EKF) algorithm [5]. The concept is based on a kinematic relation which states that on a rigid body the acceleration of any point can be determined if the angular velocity, angular acceleration and linear acceleration of another point of the body are known. A serial kinematic chain of rigid bodies mimicking human lower extremities with one wearable inertial and magnetic measurement unit (IMU) placed on each segment is presented in Fig. 1a. Quantities marked with subscript j correspond to the j-th segment. Vectors I\u03c9, Ia and IB refer to angular velocity, linear acceleration and magnetic field measured by IMU sensor, respectively. Vector Ea1 denotes the linear acceleration of the center point of the joint connecting the j-th segment with the previous j \u2212 1 segment. Vector Ea2 denotes the linear acceleration of the center point of the joint between the j-th segment with the following j + 1 segment. Accelerations Ea1 and Ea2 are expressed in the Earth\u2019s coordinate frame which is defined by the gravity vector (EY axis) and a normal vector to a plane described with gravity and magnetic vectors (EZ axis)",
            "With a recursive procedure processing segment by segment and known either acceleration of the mounting point of the first segment or it\u2019s orientation, the orientations of any number of segments in a kinematic chain can be determined. For assessing the kinematic parameters in human walking a seven segment model of human body was utilized incorporating left and right foot, shank, thigh, and head\u2013 arm\u2013trunk (HAT) segment. A recursive algorithm composed of seven EKFs is implemented for the estimation of the individual segments orientation.When the right foot is in the stance phase the recursive algorithm calculate the orientations in the rightleft (RL) direction as illustrated in Fig. 1b: (1) right foot, (2) right shank, (3) right thigh, (4) HAT, (5) left thigh, (6) left shank, and (7) left foot. In situation when the left foot is in the stance phase, the direction of the recursive calculation is reversed (LR direction). The insole data are used to determine standing and swinging leg. While the first segment is in contactwith the floor, the position of the contact center point (foot center of pressure\u2014COP), measured by instrumented shoe insoles, is used to determine the vector Sr1,1 pointing from the sole to the IMU"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003880_b978-0-12-409547-2.11551-3-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003880_b978-0-12-409547-2.11551-3-Figure1-1.png",
        "caption": "Figure 1 Effect of buffer pH, at constant ionic strength, on the mobility of the electroosmotic flow (indicated as Uosm) in 75 mm i.d. Pyrex, 75 mm i.d. fused silica, and 120 mm i.d. Teflon capillaries, all 50 cm long. Electroosmotic flow was measured by using phenol as a neutral marker with detection at 280 nm. Reprinted from Lukacs, K. D., and Jorgenson, J. W. (1985) Capillary zone electrophoresis: effect of physical parameters on separation efficiency and quantitation. Journal of High Resolution Chromatography, 8: 407\u2013411. http://dx.doi.org/10.1002/jhrc.1240080810. With permission from Wiley-Interscience.",
        "texts": [],
        "surrounding_texts": [
            "Buffered background electrolytes (BGE) are essential in capillary electrophoresis technique. Thus, when a buffer is placed inside the capillary, its inner surface acquires a charge. This is due to the ionization of the silanol groups present in the silica capillary or to the adsorption of the buffer ions onto the capillary walls. The ionized capillary wall is partially neutralized by a layer of strongly held counter-ions, which forms the fixed layer. To complete the neutralization of the capillary wall another counter-ion layer not tightly held, the mobile layer, is created performing both a diffuse double layer of ions. When an electric field is applied, the outer layer of solvated ions, usually cations, is pulled towards the negative charged cathode causing the electroosmotic flow. Between the two layers there is a plane and an electrical imbalance appears in this plane, which is the potential difference across the layers, named zeta potential. This potential, z, is defined by z\u00bc 4pde e [1] where d is the thickness of the diffuse double layer, e is the charge per unit surface area and e is the dielectric constant of the buffer. The electroosmotic mobility of the buffer, mEOF, is given by mEOF \u00bc ez 4pZ [2] where Z is the buffer viscosity. Note that the electroosmotic mobility is independent of the applied electric field and depends solely on buffer characteristics, that is, dielectric constant, viscosity, pH, and concentration, since the last ones influence the zeta potential. The effect of the buffer pH on the electroosmotic flow on a variety of capillaries is given in Figure 1which shows that the higher the pH, the higher the ionization of the capillary wall and, then, the z and mEOF values. Moreover, an increase of ionic strength or buffer concentration results in a decrease of the electroosmotic flow since it lowers the zeta potential. This effect is depicted in Figure 2 and Figure 3 for several common buffers.1,2"
        ]
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.2-1.png",
        "caption": "Fig. 4.2 Examples of application of multibody systems: a automobile vehicle model; b human biomechanical model; c four-bar mechanism with two revolute clearance joints; d shaping machine",
        "texts": [
            " A multibody modeling and formulation can be used to study the kinematic and dynamic motion characteristics of a wide variety of systems in a large number of engineering fields of application. Multibody systems can range from very simple to highly complex. There is no doubt that multibody systems are ubiquitous in engineering and research activities, such as robotics (Zhu et al. 2006), heavy machinery (Seabra et al. 2007), automobile systems and components (Ambr\u00f3sio and Ver\u00edssimo 2009), biomechanics (Silva and Ambr\u00f3sio 2002), mechanisms (Flores et al. 2008), railway vehicles (Pombo and Ambr\u00f3sio 2008), space systems (Ambr\u00f3sio et al. 2007), just to mention a few. Figure 4.2 shows some multibody system examples of application, which result from the association of structural and mechanical subsystems with the purpose to transmit or transform a given motion. In a simple way, multibody systems methodologies include the following two phases: (i) development of mathematical models of systems and (ii) implementation of computational procedures to perform the simulation, analysis and optimization of the global motion produced. Prior to establishment of the equations of motion that govern the dynamic behavior of multibody systems, it is first necessary to select the way to describe them"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001292_0954406215582014-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001292_0954406215582014-Figure1-1.png",
        "caption": "Figure 1. Schematic configuration of double contacts between two half planes (1) and (2).",
        "texts": [
            "5 for single plane contact problems is extended to analyze a generic quasi-static symmetric double contact problem between two elastically similar half planes, under the constant normal and oscillatory tangential loading, by utilizing the Ja\u0308ger\u2013Ciavarella principle. Furthermore, the consistency condition in symmetric double contacts is derived for whichever of the pressure and corrective shear functions. Next, this condition is utilized to determine the extent of the contact and stick zones in the symmetric indentation of a flat surface by two rigidly interconnected wedge-shaped punches (which has not been solved analytically yet) and the values of the pressure and shear functions are calculated in any nonsingular point of the contact zones. In Figure 1, incomplete symmetric double contacts between two half planes (1) and (2), under the constant normal and oscillatory tangential loading, are illustrated schematically. The upper half plane is considered as the indenter with a concavity on the lower half plane which was flat before loading. In the first stage of loading, the normal force P (that coincides with the y-axis which is the axis of symmetry of the indenter) is applied and increased monotonically to P0. Meanwhile, no rigid body motion is assumed for the two half planes and they are continually in the statistical equilibrium state"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002802_iet-cta.2016.0970-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002802_iet-cta.2016.0970-Figure2-1.png",
        "caption": "Fig. 2 Consensus architecture: \ufffd\ufffd(\ufffd), \ufffd\ufffd(\ufffd) and \ufffd\ufffd(\ufffd) are the plant, the local feedback control, and the pre-filter of the agent \ufffd; and \ufffd\ufffd\ufffd and \ufffd\ufffd\ufffd are the delay and the weight of the edge \ufffd\ufffd\ufffd = (\ufffd\ufffd, \ufffd\ufffd), with \ufffd\ufffd\ufffd = 0, if \ufffd\ufffd\ufffd \u2209 \u2130",
        "texts": [
            " We also define the in-degree of \ufffd\ufffd to be \ufffd\ufffd\ufffd(\ufffd) := \ufffd\ufffd\ufffd(\ufffd). Note that \ufffd\ufffd\ufffd(\ufffd) \u2208 {0, 1}, and, further, if \ufffd\ufffd\ufffd(\ufffd) = 0, \ufffd\ufffd\ufffd(\ufffd) = 0, \u2200\ufffd = 1, 2, \u2026,\ufffd (with\ufffd\ufffd\ufffd > 0). Now, suppose that each agent has its own variable of interest\ufffd\ufffd(\ufffd) \u2208 \u211d. What we want to achieve here is consensus, i.e.lim\ufffd \u2192 \u221e |\ufffd\ufffd(\ufffd)\u2212 \ufffd\ufffd(\ufffd) | = 0,with \ufffd\ufffd(\ufffd), \ufffd\ufffd(\ufffd) \u2192 \ufffd (3)\u2200\ufffd, \ufffd \u2208 {1, 2, \u2026,\ufffd}, where \ufffd \u2208 \u211d is generally a non-zero consensus value depending on initial condition. For this, we propose the twodegree-of-freedom consensus architecture as shown in Fig. 2, where the open-loop dynamics of each agent \ufffd\ufffd(\ufffd) is first stabilised by a local feedback control \ufffd\ufffd(\ufffd) and then shaped by a pre-filter\ufffd\ufffd(\ufffd) so as to attain the consensus condition as specified in Theorem 1. Here, all \ufffd\ufffd(\ufffd),\ufffd\ufffd(\ufffd),\ufffd\ufffd(\ufffd) are real rational transfer functions. This architecture turns out to allow for robust consensus synthesis for general linear agent dynamics \ufffd\ufffd(\ufffd), as long as it does not possess zeros at \ufffd = 0 \u2013 see Lemma 1. The closed-loop dynamics of the agent \ufffd is then given by: \ufffd\ufffd(\ufffd) = \ufffd\ufffd(\ufffd)\ufffd\ufffd(\ufffd) \ufffd\ufffd(\ufffd) \u2211\ufffd \u2208 \ufffd\ufffd\ufffd\ufffd\ufffd\ufffd\u2212\ufffd\ufffd\ufffd\ufffd\ufffd\ufffd(\ufffd)\u2212 \ufffd\ufffd(\ufffd) + \u0394\ufffd\ufffd (\ufffd) where \ufffd\ufffd(\ufffd) \u2208 \u2102 is the Laplace transform of \ufffd\ufffd(\ufffd), \u0394\ufffd\ufffd (\ufffd) \u2208 \u2102 is the polynomial of \ufffd related to the initial conditions of \ufffd\ufffd, \ufffd\ufffd, and \ufffd\ufffd\ufffd and \ufffd\ufffd\ufffd are the delay and weighting factor of the link \ufffd\ufffd\ufffd, with\ufffd\ufffd\ufffd = 0 if \ufffd\ufffd\ufffd \u2209 \u2130",
            " This null-space \ufffd = [1, 1, \u2026, 1]T of\ufffd, however, is not shared by diag[1, 0, \u2026, 0], whose null-space is given by [\ufffd, 0, \u2026, 0], \ufffd \u2208 \u211c. This then implies that \ufffd \u2212 \ufffd(0) does not possess any non-trivial null-space, thereby, excluding the possibility of the closed-loop consensus dynamics (10) having characteristic roots at \ufffd = 0 (since det (\ufffd \u2212 \ufffd(0)) \u2260 0). Now that the closed-loop dynamics (10) can have characteristic roots only strictly within \u2102\u2212, we can conclude that\ufffd(\ufffd) = [\ufffd1(\ufffd); \ufffd2(\ufffd); . . . \ufffd\ufffd(\ufffd)] \u2192 0 \u2200\ufffd(0) \u2208 \u211c\ufffd. \u25a1 The pre-filter \ufffd\ufffd(\ufffd) in Fig. 2 substantially facilitates the synthesis of \ufffd\ufffd(\ufffd) satisfying the consensus conditions (6) and (7): we can first stabilise \ufffd\ufffd(\ufffd) by using \ufffd\ufffd(\ufffd) and then shape the stabilised dynamics by using \ufffd\ufffd(\ufffd) in a separate/sequential manner. Synthesising \ufffd\ufffd(\ufffd), which is both stabilising and achieving the gain condition (7) at the same time is typically much more involved, particularly when \ufffd\ufffd(\ufffd) is non-minimum phase, as the closed-loop dynamics is in general under some constraints (e.g. complementary sensitivity integral condition [22]). In fact, this prefilter \ufffd\ufffd(\ufffd) and its two-degree-of-freedom architecture in Fig. 2 allow us to grant the consensus conditions (6) and (7) as long as\ufffd\ufffd(\ufffd) has no zeros at \ufffd = 0. Lemma 2: The consensus conditions (6) and (7) in Theorem 1 can be granted by using the local feedback \ufffd\ufffd(\ufffd) and the pre-filter\ufffd\ufffd(\ufffd), if and only if the agent dynamics \ufffd\ufffd(\ufffd) does not have any zero at \ufffd = 0. Proof (Necessity): Suppose \ufffd\ufffd(\ufffd) has zeros at \ufffd = 0. Since open-loop zeros are not affected by \ufffd\ufffd(\ufffd), to attain (6), \ufffd\ufffd(\ufffd) should contain the same number of poles at \ufffd = 0, resulting in pole-zero cancellation at \ufffd = 0, which is not allowed in Theorem 1",
            " Lastly, note that the consensus control proposed in this paper is robust against any constant/unknown time-delay, and its synthesis is also completely decentralisable (i.e. each agent can design its own control without consulting others). We consider four heterogeneous agents on the first information graph as shown in Fig. 1. We impose non-uniform constant delays:\ufffd12 = 0.1 s, \ufffd23 = 0.3 s, \ufffd31 = 1.5 s, and \ufffd43 = 0.15 s. Agent 3 is a second-order system with the following dynamics:\ufffd\ufffd\ufffd\u0308\ufffd(\ufffd) + \ufffd\ufffd\ufffd\u0307\ufffd = \ufffd\ufffd(\ufffd), where \ufffd\ufffd, \ufffd\ufffd > 0 are uncertain mass and damping parameters. We then design the consensus control\ufffd\ufffd(\ufffd) := \u2212 \ufffd\ufffd(\ufffd\ufffd(\ufffd)\u2212 \u2211\ufffd \u2208 \ufffd\ufffd\ufffd\ufffd\ufffd\ufffd\ufffd(\ufffd \u2212 \ufffd\ufffd\ufffd)) so that, for Fig. 2,\ufffd\ufffd(\ufffd) = 1/(\ufffd\ufffd\ufffd2+ \ufffd\ufffd\ufffd), \ufffd\ufffd(\ufffd) = \ufffd\ufffd, \ufffd\ufffd(\ufffd) = 1 and \ufffd\ufffd(\ufffd) = \ufffd\ufffd\ufffd\ufffd\ufffd2+ \ufffd\ufffd\ufffd+ \ufffd\ufffd where, if we set \ufffd\ufffd \u2264 (\ufffd\ufffd2/4\ufffd\ufffd) (i.e. critically-damped) for all possible \ufffd\ufffd, \ufffd\ufffd, the consensus conditions (6) and (7) of Theorem 1 can be robustly ensured. Agents 2 and 4 are first-order systems with the following dynamics: \ufffd\u0307\ufffd(\ufffd) = \ufffd\ufffd(\ufffd), with the consensus control \ufffd\ufffd(\ufffd) designed by \ufffd\ufffd(\ufffd) := \u2212 \ufffd\ufffd(\ufffd\ufffd(\ufffd)\u2212 \u2211\ufffd \u2208 \ufffd\ufffd\ufffd\ufffd\ufffd\ufffd\ufffd(\ufffd \u2212 \ufffd\ufffd\ufffd)), \ufffd\ufffd > 0. Then, for Fig. 2, we have \ufffd\ufffd(\ufffd) = 1/\ufffd, \ufffd\ufffd(\ufffd) = \ufffd\ufffd, and \ufffd\ufffd(\ufffd) = 1, with \ufffd\ufffd(\ufffd) = \ufffd\ufffd\ufffd+ \ufffd\ufffd which satisfies the consensus conditions (6) and (7) in Theorem 1. Agent 1 is a non-minimum phase fourth-order system with relative-degree 3: \ufffd\ufffd(\ufffd) \u2208 \ufffd~\ufffd := \ufffd \u2212 \ufffd\ufffd(\ufffd+ 1)(\ufffd2+ \ufffd\ufffd\ufffd+ \ufffd\ufffd) (17) where \ufffd\ufffd \u2208 [2, 18] and \ufffd\ufffd \u2208 [150, 250] with \ufffd\ufffd = \ufffd\ufffd/50 (i.e. dcgain \ufffd\ufffd(0) = \u2212 1/50). First, we design a robust stable local feedback control \ufffd\ufffd(\ufffd) s.t., \ufffd\ufffd(\ufffd) := 200 \ufffd \u2212 1\ufffd(\ufffd+ 10) with which \ufffd\ufffd(\ufffd) = (\ufffd\ufffd(\ufffd)\ufffd\ufffd(\ufffd))/(1 + \ufffd\ufffd(\ufffd)\ufffd\ufffd(\ufffd)) is stable\u2200\ufffd\ufffd(\ufffd) \u2208 \ufffd~\ufffd with \ufffd\ufffd(0) = 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002948_gt2016-57458-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002948_gt2016-57458-Figure9-1.png",
        "caption": "Figure 9 DFP inner support structure overview.",
        "texts": [
            " The ribs are designed such, that the projected area of the outer support structures would stay within the projected contour of the DFP on the substrate plate (see Figure 8 for details). A number of diamond cut outs was implemented in the ribs to save material and make the supports structure not too heavy. During further investigation of the DFP outer shape, another SLM critical edge was identified. The edge was approximately 15mm wide and 150mm long. For overhang regions with such extensions additional inner supports were necessary. The inner support structure (see Figure 9) consists of two families of narrow supports, which are crossing each other. In the area of the critical edge a number of tangential features were added to secure the smooth transition of the single SLM layers. 4 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89513/ on 04/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use During the first SLM production trial of full size DFP the process was stopped at \u00be height of the panel (see Figure 10)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003173_urai.2016.7734120-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003173_urai.2016.7734120-Figure7-1.png",
        "caption": "Figure 7 shows the mechanical fault simulator and the position of gearbox in the mechanical fault simulator.",
        "texts": [
            " According to the vibration of the gear box model, assuming that the original signal as and using simulated signals to verify. The sampling points is 2000,the DC offset is 0 and the input phase is O. , Ij\\{V\\{\\/\\{\\r\ufffdJ\\{J\\{V\\(\\/\\{\\/\\{\\/1 \u00b7:t\ufffd ,:, .:, u u \u2022 .:, .:. .:. 1 Fig.6 Decomposition results of the simulated signal. Using the EMD program written in LABVIEW to decomposition the simulated signal, the original signal and decomposition results are shown in fig.6.The decomposition results shows that the EMD program can decompose signal correctly. 5 Experimental analysis 5.1 Experimental setup Fig.7 Position of gearbox in the Machinery Fault Simulator. The experiment was conducted on the Machinery Fault Simulator (MFS) launched by Spectra Quest Company. As it shown in figure 8 (a) that the ratio of the gear transmission is 1.5: LAnd 8 (b) identifies the two spur bevel gears in the gearbox. Vibration Data is collected by triaxial accelerometer installed on the top of gearbox. Its sampling frequency is 4096, and sampling length is lOs, and rotational speed of the input axis is 30r/menthe transmission ratio is 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure2.22-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure2.22-1.png",
        "caption": "Fig. 2.22 Conformity of points for positive and negative load",
        "texts": [
            "6) of load power in the similar relative form P \u00bc PL PSC 0 \u00bc KV \u00f0KV \u00de2; \u00f02:49\u00de where PSC 0 is the maximum power of the voltage source V0 for SC regime. Therefore, the task of equal regimes does not cause a problem; that is simply corresponding equality of values KV ; I;P. But a deeper analysis will be useful, which allows generalizing the justification of the equality of regimes and will be used for considering a more complex case, the efficiency of two-ports. 2.2 Volt\u2013Ampere Characteristics of an Active Two-Pole with a Variable Element 47 2.3.1 Symmetry of Consumption and Return of Power Let us consider load straight line (2.48) in Fig. 2.22. In the first quadrant, a positive load consumes energy; there is the maximum load power point P\u00fe LM for RL \u00bc Ri. At SC and OC regime points, RL \u00bc 0;1, load power (2.49) is equal to zero, as shown in Fig. 2.23. We remind if a negative load returns energy into the voltage source V0, then the load voltage VL\\0 and VL [V0. In this case, the load resistance RL\\0; the load power increases. In the final analysis, the load power P LM \u00bc 1 for the resistance RL \u00bc Ri. Therefore, the load straight line is closed and the parabola is a closed oval curve too, which concerns the infinitely remote straight line TP LM at the point P LM "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002437_s1068798x16040195-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002437_s1068798x16040195-Figure1-1.png",
        "caption": "Fig. 1. Pressure regulator (a) and toroidal shutoff valve (b).",
        "texts": [
            "70 Reliable operation of hydraulic components in drives and automatic systems based on traditional technology and high quality structural materials depends on high precision and machining quality of the contacting surfaces, which are very sensitive to contamination of the working fluid. Elevated performance requirements on such com ponents may be met if the basic elements of the regu lators are made of the elastic materials traditionally employed in sealing. hydraulic components were con sidered in [1\u20134]. An elastic element may be used, for example, in the second (primary) stage of the pressure regulator in Fig. 1a [2]. We now consider regulator operation. If the pres sure in the system is less than the pressure setting of the control (auxiliary) valve 7 (in the first stage), it closes, and the pressure in chambers 1 and 2 is the same. The primary shutoff valve 3, which is toroidal (Fig. 1b), is pressed into its saddle, thanks to the difference in the surface areas on which the constant pressure acts. The input chamber 1 is isolated from the output chamber 4 in this case. With increase in pressure in the system, control valve 7 opens, and a pressure difference appears at choke 5: the pressure in chamber 1 is greater than that in chamber 6. As a result, shutoff valve 3 moves axially to the left, thereby connecting input chamber 1 and output chamber 4. The equation for the forces in the regulator corre sponds to the recommendations of [5] (1) where ps is the supply pressure; pco is the pressure ahead of valves 3 and 7; Phd is the hydrodynamic force; Pfr is the frictional force, m is the mass of valve 3; h is its dis placement; F is the cross sectional area of toroidal valve 3; the factor a takes account of the cross sectional area of the toroid in the zone where pressure ps acts"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000622_aim.2015.7222655-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000622_aim.2015.7222655-Figure6-1.png",
        "caption": "Fig. 6 3D plot showing the locations of discontinuities",
        "texts": [
            " 3) as follows: Step 1 (Fig. 4): From the gray-scale (or pseudo color) infrared image, obtain the potential (temperature) and streamline (gradient) of the rake face and its vicinity, and determine the peak temperature (at which / 0T s   ). Step 2 (Fig. 5): Check for discontinuities in T and/or T/s along the streamlines. A discontinuity in T but not in T/s indicates that the tool and chip are in contact. If both T and T/s are discontinuous, the tool and chip are in different planes without contact. Step 3 (Fig. 6): Once the discontinuities are marked, the chips (regions A and B) that obtrude the tool insert can be segmented and removed, where region A corresponds to the tool-workpiece contact area but obtruded by the chip. Step 4 (Fig. 7): The missing temperature of the tool insert 0 1 2 3 4 5 6 7 8 200 300 400 500 A B C 5 10 s=0 T (\u00b0 C) s(mm) A B C IR data Curve-fit Reconstructed temperature T (K) B Fig.7 Illustration of temperature reconstruction of the tool insert Experiments were performed to investigate the thermal and surface properties on the temperature distribution of a typical titanium alloy workpiece, and to identify key parameters that significantly influence its performance during machining",
            " 8, a 1mm-diameter heat wind-gun was employed as a controllable heat source positioned at 100mm from the rotating center. Its outlet is perpendicular to the WP surface) and at 3 mm from the WP surface on which a thermocouple was placed. The opposite side was measured by a temperature probe (HASCO/Z2511, 0.1C resolution) for verifying IR camera recordings. As compared in Fig. 8 where data were obtained with 4 different air-speeds, the IR error is within 2.5% of that measured by the temperature probe. A steep temperature gradient cross the 5mm thickness can be seen in Fig. 6, which increases with temperature. Transient and steady-state responses of TC4 workpiece Figure 9(a) shows the setup for Exp. 1a where a TC4 disk (300mm-dia.) was held an aluminum holder. The transient response of the peak temperature for the TC4 disk is given in Fig. 10. To provide an intuitively comparison, the transient response experiment was repeated on a same-size Aluminum disk with results shown in Fig. 11. TC4 has a much larger cooling time-constant (55s) than that of aluminum (15s). Unlike aluminum which dissipates heat rapidly (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003994_ee.1935.6539669-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003994_ee.1935.6539669-Figure1-1.png",
        "caption": "Fig. 1 . Cy l in - d r i c a l r o t o r s y n c h r o n o u s generator con nected to an infinite bus",
        "texts": [
            " (In reference 3, at the end of the paper, the effects on the saturation of the change in field leakage flux under load conditions are approxi mately accounted for by using, instead of the open-circuit saturation curve, a family of load saturation curves calculated with the appro priate values of field leakage under load conditions. If deemed ad visable, this refinement may be introduced. It is omitted in this paper in order to simplify the presentation of more essential facts.) SATURATED SYNCHRONOUS MACHINE Consider a saturated cylindrical-rotor synchronous machine operating under balanced steady-state con ditions, connected through an impedance Ze of re sistance re and reactance xe to an infinite bus of voltage V, as shown in figure 1. On the basis of the foregoing assumptions, the well-known vector dia gram shown by the solid lines of figure 1 can be drawn,3'6 in which i* = armature current Vt = terminal voltage of the machine r\u201e = armature effective resistance xa = armature leakage reactance Ea = voltage generated by the resultant air gap flux E* = voltage which would be generated by the resultant magneto motive force R if the magnetic circuit were not saturated Ef = voltage which would be generated by the field excitation / / if the magnetic circuit were not saturated Xd = unsaturated value of the synchronous reactance xm = Xd \u2014 xa = reactance equivalent to the armature reaction when there is no saturation The ( i ) Any consistent system of units may be used, voltages Ef and EJ are shown in figure 2 . Let k - Ea'/Ea Then k is a measure of the saturation of the machine. (Except for the manner in which field leakage is treated, k as defined above is the same as the satura tion factor k of reference 3 at the end of the paper.) Figure 4 shows a typical curve of k as a function of Ea. By applying equation 1 to figure 1, it can be seen that the saturated value of the synchronous reactance is X, = Xa + Xm/k ( 2 ) and that the saturated value of the excitation voltage is E = Ef/k ( 3 ) These results agree with those obtained by Putman.5 Since k is a function of Eay both x, and E are func tions of Ea. Figure 4 shows a typical curve of x9 as a function of Ea. Equations 2 and 3 are derived directly from the fundamental vector diagram of figure 1, and hence take into account correctly the most important effects of saturation. This is an important improve ment over any method of empirically adjusting the synchronous reactance. 2 It can readily be seen that the voltage E in figure 2 is equal to the voltage E given by equation 3 . (In figure 2 , EJE = EJ/Ea = k, or E = Ef/k). It should be noted that the excitation voltage E is not equal to the voltage E' read from the open-circuit saturation curve, figure 2 . The voltage E' would appear at the machine terminals at no-load if the field current were maintained constant. Hence E' is frequently used as the excitation voltage. How ever, the removal of the load would change the satu ration of the machine. The voltage E would appear at the machine terminals if the load were removed and the field current and saturation maintained the same as under load conditions. Hence equation 3 determines the excitation voltage at the correct satu ration of the machine under load conditions. POWER-ANGLE CHARACTERISTICS Consider the system shown in figure 1: As is well known, the air gap power of the machine Pa is given by EV . , , , E*r P a = ~z s i n ( - a ) + Z*~ ( 4 ) and the power Pb delivered to the infinite bus is given by P 6 = ^ s i n ( S + \u00ab ) - ( 5 ) where r = total series resistance = re + ra x = total series reactance, saturated value = xe -f- x, Z \u2014 total impedance = V r 2 - j - x2 a = t a n - 1 (r/x) 5 = load angle of the machine with respect to the infinite bus, i. e., the angle between the vector voltages E and V, positive when E leads V",
            " The following cal culation for one assumed value of 8 illustrates the method: For 5 = 60.0 degrees, {kZEa)2 = 2.57, from Equation 8a Hence kZ = 1.54, a = 5.2 degrees, Z 2 = 1.55, from figure 5 Hence, from equations 6 and 7 at 5 = 60.0 degrees Pa = 0.980 Pb = 0.876 Figure 6 also shows a comparison of calculated with test results. The agreement is well within the accu racy required in most engineering problems, the cal culated values of power being, at the worst, about 0.02 per unit below the test results. CALCULATION OF M A X I M U M POWER Consider the system of figure 1. If in equation 6, (kZ) and a were constants, then under specified con ditions of constant excitation and infinite bus volt- 1 1.0 P\u00a30 .6 | < | O 0 . 4 \u00a3 5 0 - 2 > / \\ \\ / \\ / i / / k / 0 2.0 'i 1.6 z o 4 1.0 Fig. 4 . Satu ration factor and saturated s y n c h r o n o u s r e a c t a n c e as (unctions of the air sap voltage 0.2 0.4 0.6 0.8 1.0 PER-UNIT AIR-GAP VOLTAGE, 1.2 Ea age, the air-gap power-angle characteristic would be a displaced sinusoid, and the maximum value of Pa would obviously be EfV Ef*r \u00b1 P a m a x ",
            " However, in all practical cases, near its maximum value, the powerangle characteristic is approximately a displaced sinusoid and very little error will result from as suming that at (b \u2014 a) = =\u00b1=90 degrees, the corre sponding value of Pa given by equation 9 is the maximum value. The error will not be over one per cent if the load angle at which the air gap power actually reaches its maximum value is within about 8 degrees of (5 \u2014 a) = 90 degrees. In equation 9, (kZ) is dependent upon the saturation. Its value at pull-out can be determined by substituting 5 = (=\u00b1=90 degrees + a) in equation 8, and referring to the auxiliary curve of (kZ) as a function of (kZEa)2, similar to that shown in figure 5. vSimilarly, from equation 7, the maximum power which the machine of figure 1 can deliver to the re ceiver bus under specified conditions of constant excitation and receiver bus voltage is, very nearly, jy EfV V*r (10) at a load angle 5 = 90 degrees \u2014 a. The values of (kZ) and Z 2 to be used in equation 10 can be deter mined by substituting 5 = 90 degrees \u2014 a in equa tion 8 and referring to the auxiliary curves of figure 5. NUMERICAL EXAMPLE 2 Consider the system of example 1, shown in figure 1. Let it be required to calculate the pull-out power of the generator at a constant voltage of the receiver bus, V = 1.00, and a constant excitation, Ef = 1.61 (as given in \"Example 1\"). Pull-out will occur approximately when 8 = (90 degrees + a). The angle a is a variable dependent Fig. 5. curves pie 1 Auxi l ia ry for cxam- 1.2 1.0 \\ ^ oc^ kZ en hJ a I 5 8 UJ _ i O 1 z MARCH 1935 3 0 3 PE R U N IT _ kZ A N D _ Z 2 ^ 1 <L 3 2 4 5 PER UNIT ( k Z E a ) Test p o i n t s s h o w n b y small c i r c l e s upon saturation, and hence its exact value is not known",
            " Table I shows a comparison of the maximum values of the air gap power calculated for several representa tive cases by the following methods of calculation: Method A. From peak value of the air-gap power-angle characteris tic calculated as shown in example 1. Method B. Calculated by equation 9, as shown in example 2. Method C. Neglecting saturation entirely. Method D. Adjusted synchronous reactance, as described in refer ence 1 of the list at the end of the paper. Method E. Adjusted synchronous reactance, as described in refer ence 2. The results are shown for the following cases: Case 1. System described in example 1. Case 2. System of figure 1. Machine of example 1. re = 0.482, xe = 0.661, V = 1.00, Ef = 1.70 (Large resistance). Case 3. System of figure 1. Machine of example 1. re = 0, = 0.500, V = 1.00, Ef = 2.50 (Large field current). Case 4. Machine of example 1 directly connected to an infinite bus. V = 1.04, Ef = 1.74. Method A is based directly upon the fundamental vector diagram of figure 1, which is known to give reasonably accurate results, as shown for case 1 by the comparison of calculated and test results in figure 6. Hence the results obtained by method A can be accepted as substantially correct, and these results used as a standard of comparison in judging Table II\u2014Comparison of Results Obta ined b y Various Methods of Calculating Synchronizing Power the accuracy of the other methods. On this basis, the errors of the other methods are shown in table I. Table II shows a comparison, of the values of the synchronizing power, dPa/d8, for the system of example 1 under the following load conditions: Generator delivering its rated kilovoltamperes with its excitation adjusted so that its terminal voltage Vt = 1",
            " Reference 3. SUMMARY It is important to take into account correctly the effects of saturation in calculating the performance of synchronous machines. It is shown in this paper that these effects can be correctly and conveniently taken into account by means of equations 2 and 3 as expressions for the saturated synchronous reactance and excitation voltage of a cylindrical rotor syn chronous machine operating under balanced steadystate conditions. A p p e n d i x A \u2014 D e r i v a t i o n of Equation 8 From figure 1: _ _ _ - E \u2014 V Ea = V + I Zc and I = - in which all quantities are expressed as vectors. Z = r + jx = (re + ra) + j(xt + xa + xjk) Zc = (r , + r\u00ab) + j{xe + *\u00ab) Hence p _ 7/ , & ~ vTZq = \u00a3 Z C + V(Z - Ze) z z Since E = ~Ef/k (equation 3) , and (Z - Zc) = jxm/k, ZEa = ^Zc+jV\u00b0% k k Hence kZE* = EfZc + jVxm Method dPa/db % error F 0.653 0 G 0.647 - 1 H 0.741 + 1 3 / 0.549 - 1 6 J 0.749 + 1 5 (ID Let V be the reference vector. Then in polar form, V = V/Q de grees and Ef = Ef/8t where the symbols V and Ef (without the bars above them) mean the magnitudes of the voltages"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000392_978-1-84996-432-6_26-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000392_978-1-84996-432-6_26-Figure2-1.png",
        "caption": "Fig. 2. Schematic diagram of the experimental set-up",
        "texts": [
            " A special type calibrated probe, made indigenously, is used to measure air pressure at various points radially outwards from the wheel peripheral surface (R-axis) as shown in Fig. 1. The probe of 1mm outer diameter is set at 0.1mm distance from the middle position of wheel periphery. U-tube inclined manometer with water as manometric fluid and inclination angle of 15o is used to measure the air pressure. A pneumatic nozzle of 4mm internal diameter is placed on wheel face at different polar co-ordinates (r, \u03b8) (Fig. 2). The polar angles (\u03b8) considered are 30o, 45o and 60o. The pneumatic nozzle is placed 10mm away from the middle position of the wheel periphery, so that curve radius, r becomes 110mm. The pneumatic nozzle is pivoted at P, and swivelled at an angle \u03b1. The axis of pivot is parallel to the grinding wheel spindle. The swivel angle (\u03b1) is taken as 0o, 30o, 50o, 70o and 90o. Through this nozzle, compressed air is supplied to block the stiff air layer before the probe. Compressed air is applied with pressure varying from 100 mm to 1600 mm of water",
            " Substantial amount of reduction in air pressure at the middle position of the periphery of grinding wheel is achieved by using pneumatic barrier. From Fig. 3, the maximum reduction of air pressure at peripheral middle position of grinding wheel is observed at a swivel angle (\u03b1) of 50o with a pneumatic pressure of 800mm of water. From Fig. 4, it is seen that the maximum effect of pneumatic barrier is near about 53% at a pneumatic pressure of 800mm of water and at a \u03b1 of 50o. Face of the pneumatic nozzle is pivoted at point P. Hence, line of action of pneumatic pressure is at a swivel angle (\u03b1) with the normal OP (Fig. 2). Line of action of pneumatic pressure is tangential at \u03b1= 65o. At an \u03b1= 0o, pneumatic air strikes the wheel perpendicularly. Therefore, the effect of blocking the tangentially flowing air, which is generated around grinding wheel, becomes less. At \u03b1= 30o and 50o, compressed air strikes the wheel at polar angles (\u03b8 +3o) and (\u03b8 +7o) respectively. At these swivel angles, effective penetration of stiff air layer by the pneumatic barrier is expectedly observed. However, increasing the swivel angle (\u03b1) beyond 500 causes large increase in the distance of striking of pneumatic air with wheel from the probe, and hence, the effect of blocking air layer around grinding wheel becomes less"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001306_s1063454114040086-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001306_s1063454114040086-Figure1-1.png",
        "caption": "Fig. 1. A two pendulum cart system.",
        "texts": [
            " alized Gauss principle, because in this case it is only required to increase the order of the principle to be used to a value agreeing with the number of given boundary value conditions. Some numerical results are given. 2. STATEMENT OF THE PROBLEM To illustrate the theories under consideration we state the following specific problem. Assume that on a lift crane trolley of mass M, which rides on horizontal rails along the x axis, there are fixed ropes of lengths, respectively, l1 and l2. The loads of the masses m1 and m2 are suspended at the ends of these ropes (see Fig. 1). It is required, for a fixed time , by choosing and applying a horizontal force F(t) to a trolley, to transfer the suspended loads to a given distance a, from a state of rest to a state of rest. The motion equations of this system are written, for small oscillations, as follows (g is the acceleration of gravity): (2.1) To stop oscillation when the system is stopped, considering that the system is at rest ab initio, the con trol force should be such that the following boundary value conditions are satisfied: (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003046_j.ifacol.2016.07.967-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003046_j.ifacol.2016.07.967-Figure1-1.png",
        "caption": "Fig. 1. One-rotor vibration unit.",
        "texts": [
            " In this paper we have made design and numerical study of time-varying observer operability, reducing the rate of vertical movement of the supporting body (platform) of a single rotor vibration unit, oscillating in a vertical plane. Efficiency of the developed algorithms for passage through resonance zone based on the proposed observer is analysed. To study the efficiency of control algorithms for the main modes of vibration units the mechatronic vibration setups with one unbalanced rotor were developed (Blekhman et al., 1999). The schematics of one-rotor unit, whose supporting body is oscillating in a vertical plane 0xy is shown in Fig. 1. Here, a coordinate system associated with the supporting body is designated as 0'\u0445'y'. The coordinates of the center of the supporting body \u2013 point 0' in the fixed coordinate system are denoted as {\u0445\u0441, y\u0441}. The rotor is located in the center of the supporting body. The supporting body is reinforced by elastic supports \u2013 springs located at \u0445ni = \u0430. The rotation angle of the supporting body denoted as \u03c6sb. In addition, the following notation is used: \u03c6 \u2013 rotation angle of an unbalanced rotor, measured from the axis 0x' counterclockwise; msb, m \u2013 the masses of the supporting body and the rotor; J \u2013 moment of inertia of the rotor; \u03c1 \u2013 eccentricity of the rotor; \u044101, \u044102 \u2013 longitudinal and lateral rigidity of the springs; g \u2013 acceleration of gravity; kc \u2013 viscous friction coefficient of the vibroactuator bearings;  \u2013 total axial damping springs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003942_3223384-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003942_3223384-Figure4-1.png",
        "caption": "FIG. 4. Spherical micromanipulator with a radial type reduction of lateral movement and viscous reduction of movement for longitudinal displacement of the handle.",
        "texts": [
            " It can be used even in work with the hanging-drop method. However for some operations it is too slow. For example, if we want an opening in the bony wall around a nerve and we try to take away small bone chips, the carving movements are often too slow for a good cut in the bone. For this purpose, what seems to be indicated is to reduce the lateral movements by a lever system instead of by viscosity. 2:2 X::: :m::i::::- FIG. 6. Spherical micromanipulator with lever type longitudinal displacement. Several models were constructed and the type shown in Figure 4 proved to be the simplest and most reliable. In this model the displacement along the longitudinal axis of the handle is controlled by viscous forces in the same way as in Figure 2. For the lateral movements tube 3 is attached to sphere 11 instead of a disc. The radius of this sphere is slightly larger than the distance between the tip of the instrument and the surface of the sphere. Therefore, an upward displacement of handle 7 will produce a similar, but greatly reduced, displacement of the tip of the instrument",
            " If the handle shows no resistance to movement, the position of the hand is controlled only by the receptors of the arm and the involuntary displacements of the hand are quite large. But if the handle has some stiffness, displacements of the hand produce stimuli on the tips of the fingers which in turn tend to steady the position of the hand. For some special experiments, as on the blood arterioles, maximal speed is necessary. Consequently all viscous forces must be avoided. For this purpose the lever system, shown in Figure 5, was developed to fit inside tube 3 of Figure 4. The system consists of a tube 17 in which wedge 18 glides. This wedge rotates lever 19 around its pivot 20 so that the short arm of the lever displaces the instrument fixed at the tip of 21 in the longitudinal direction. Spring 23 is used for return motion and to avoid backlash. Figure 6 is a photograph of this micromanipulator. The handles of the operative tools shown in Figure 4 are made light enough in weight so that gravitational forces will not cause them to slide downward when they are released. However, because of its increased weight, the handle shown in Figure 5 has a tendency to slide downward very slowly. To prevent this, the device shown in Figure 6 has been provided to lock it in position. In the locked position, spring 26 presses fork 25 against disc 13. Raising handle 24 lifts the spring and releases the lock. The coarse adjustment of the micromanipulator is obtained by permanently attaching handle 12 to the tube of the microscope objective in such a way that the tip of the operating instrument moves in the middle of the focused field of view of the microscope"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002425_s11771-014-2021-5-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002425_s11771-014-2021-5-Figure5-1.png",
        "caption": "Fig. 5 Tooth profile of gear slotting cutter",
        "texts": [
            " (5) can be rewritten as 1 1 1 1 1 1 1 1 1 1 cos sin sin cos 1 1 o i i i i i o i i i i i x x y y x y                        (6) Supposing that k1i is the slope of the cubic spline curve, which is reconstructed by the discrete points using the curve fitting method, the tangent vector in the coordinate system  (O, x, y) of an arbitrary point on the cosine tooth profile, after which rotating an angle \u03c61i clockwise can be represented as 1 1 1 1 1 01 1 1 1 1 1 cos sin sin cos 1 11 o xi i i i o yi i i i i k k k k k                                   M (7) 3.2.1 Mathematical model of gear slotting cutter As shown in Fig. 5, parametric equation of tooth profile of the involute gear slotting cutter in the coordinate system  2(O2, x2, y2) can be expressed as 2 b 2 b (cos sin ) (sin cos ) x r y r            (8) where rb denotes the radius of the base circle,  represents the generating angle of the involute tooth profile, \u03b1k is the pressure angle of point K, \u03b4 is the parametric variable, which can be then expressed as k    As shown in Fig. 3, the equation of tooth profile of the involute gear slotting cutter in the fixed coordinate system 2(O2, x, y) is represented by 2 2 2 02 1 2 2 2 2( ) ( ) 1 1 o o i i x x y y                    M A A (9) where M02 represents the coordinate transformation matrix from coordinate system 2 to , a is the initial distance between the gear blank and gear slotting cutter, 1 2( )iA and 2 2( )iA denote the rotation and translation matrices from coordinate system 2 to \u03a3\u2032, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000584_itec.2015.7165740-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000584_itec.2015.7165740-Figure2-1.png",
        "caption": "Fig. 2. Induction machine single-phase excitation scheme",
        "texts": [
            " (3) Here, the subscript s in the stator voltages and currents is removed to simplify the expression, and the elements in the impedance matrix are S(()2G\" R=R,+--1\\ (4) x = (()Ls where, X is the stator reactance. When (3) is compared to the original model in (1), the order of the impedance matrix is reduced from 4 to 2. With the same assumption, the input impedance (2) is approximated by (5). (5) B. 1M Reduced-order Model under Single-phase Test The single-phase excitation method can be used to replace the conventional locked rotor test [13]. The connection diagram for single-phase excitation is shown in Fig. 2. The forward component model of the single-phase excited induction machine can also be represented by (1) [12]. Because the motor is excited by single phase current, there is no rotating magnetic field, and the electromagnetic torque is zero; therefore, the rotor is not rotating at all, and the slip equals to unity (s = 1). The traditional locked rotor test is usually applied under rated frequency, such that 1\\ \u00ab UJ\ufffd, is usually assumed and the magnetizing inductance branch can be neglected [22]",
            " The frequency reference OJ and per unit voltage reference (vdPu' vqpJ ' can be adjusted to operate the machine at different speed and saturation levels, such that the stator inductance can be measured under different conditions. B. Single-phase excitation test The single-phase excitation test is employed to measure the total resistance \ufffd + (Gn / G)R, and total leakage inductance O'Ls ' The control diagram is illustrated in Fig. 5. In order to achieve the single-phase excitation scheme illustrated in Fig. 2, the 3-phase SVPWM modulation is replaced by SPWM modulation. a-axis component of the inverse Park transformation output is chosen to be the reference of phase A, and the reference voltages of phase B and phase C are the negative of phase A, as described in (14). va = va = v dplI cosB- vqpll sinO Vb = ve = -Va (14) When the machine is excited in single-phase mode (phase B and C are paralleled), considering the 3-phase windings are symmetric, the voltage applied on phase A is 2/3 of the terminal voltage UAB"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000750_s11668-015-0006-9-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000750_s11668-015-0006-9-Figure1-1.png",
        "caption": "Fig. 1 Schematic sketch of tapered roller bearing",
        "texts": [
            " Keywords Tapered roller bearing Premature failure Untempered martensite Preload Differential pinion bearing Bearing life is dependent on rolling contact fatigue material properties, heat treatment, lubrication, and operating conditions. However, premature failure of the bearing may also be caused due to improper design, selection, handling of bearing, and installation procedures [1, 2]. Tapered roller bearings are commonly used in automotive differentials to support the pinion gears. In these bearings, roller head is in sliding contact with cone large end flange (rib) and this rib portion prevents the roller from squeezing out from the bearing during normal operation as shown in Fig. 1. Bearings used for pinion application are generally preloaded at the installation stage. The contact stress between the roller and rib depends on the geometry of the rib, roller head profile, and extent of preload [3, 4]. Adequate lubrication is necessary to avoid direct metal-to-metal contact between the contacting surfaces and consequent frictional heat. Excessive preload and/or insufficient lubrication has the potential to cause undue rise in temperature due to the increased contact stress between the rib and the roller"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001394_12.2083756-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001394_12.2083756-Figure1-1.png",
        "caption": "Figure 1. Haptic Device",
        "texts": [
            " ( ) 0 0 0 ,)sgn( , ,)sgn( , \u2265 = \u2264 \u23aa \u23a9 \u23aa \u23a8 \u23a7 + + = z z z zkzD F zkzD FzF pp n nn nfriction & & & && && & (8) where z& is the relative velocity between scalpel and tissue, nF is the sum of nonfrictional force applied to the system(which includes inertial effects and force from elasticity), pD and nD are the positive and negative value of the coulomb friction coefficients, and pk and nk are the positive and negative value of viscous friction coefficients. 3. HAPTIC DEVICE To verify the force model, the haptic device using magnetorheological fluid is utilized. The haptic device is comprise of slave robot which conducts the incision process and haptic master which controls slave robot. The haptic master is a device to control slave robot and realize reaction force as shown in figure 1 (a). To guide the position of slave robot, angles of manipulation handle is measured by encoders (Autonics E40H8) which connect to its shaft. Simultaneously, the reaction force is realized by MRF BDC which can transmit the torque generated by a DC motor (D&J IG32GM 04type) into the shaft. The BDC is connected to planetary gear and it can control the torque with magnetorheological fluid changing effective magnetic field using the solenoid coils. The torque tracking using MRF BDC is much superior compared with only DC motor approach in a view point of continuous force control. The slave robot is a device to conduct incision process and collect reaction force data as shown in figure 1 (b). The slave robot is compose of 3-axis linear robot structure. The scalpel (nopa instruments BLADE#20) of end effector with force sensor (ATI-Nano17) is operated by step motor (Autonics A4k-M245) which is linked to end effector with timing belt in horizontal direction and acme thread in vertical direction. The haptic master realizes desired reaction force from force model. To achieve this haptic application, PID control with open-loop control which can control in uncertain system condition can be applied"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.4-1.png",
        "caption": "FIGURE 6.4",
        "texts": [
            " A simplistic representation of the torsional stiffness of the body may be used (Blundell, 1990) where the vehicle body is modelled as two rigid masses, front and rear half body parts, connected by a revolute joint aligned along the longitudinal axis of the vehicle and located at the mass centre. The relative rotation of the two body masses about the axis of the revolute joint is resisted by a torsional spring with a stiffness corresponding to the torsional stiffness of the vehicle body. Typically, the value of torsional stiffness may be obtained using a finite element (FE) model of the type shown in Figure 6.4. For efficiency, symmetry has been exploited here to model only one half of the vehicle body. This requires the use of anti-symmetry constraints along Finite element model of body-in-white. the centre line of the FE model, for all nodes on the plane of geometric symmetry, to carry out the asymmetric torsion case. It is also possible to incorporate a FE representation of the vehicle body within the MBS full vehicle model. An example of this was shown in Figure 4.3 where the flexibility of a racing kart frame was included in the model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003365_phm.2016.7819880-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003365_phm.2016.7819880-Figure6-1.png",
        "caption": "Figure 6. Experiment test rig.",
        "texts": [
            " An accelerometer (with sensitivity of 100mV/g and frequency range 0-10kHz) and a shaft encoder (produced by Encoder Products Co. with 1 pulse/revolution) are used for capturing vibration and Tacho signals simultaneously. The data are captured under speed up conditions from 0 RPM to 3000 RPM within 8 seconds data is used for the following analysis. The sampling frequency was set to be 7680Hz to accommodate all interested frequency contents of this test rig. The whole set-up arrangement is shown in Fig. 6 [6]. In the experimental process, four types of faulty sun gear are considered, see in Fig. 7. In the following, for demonstration purpose, vibration data from the tooth-missing fault (Fig. 7 (d)) is used to verify the effectiveness of the signal selection scheme. The physical parameters for planetary gearbox are listed in TABLE II in which gear teeth, number of planet gear and transmission ratio are given and calculated. In the experimental set-up, the ring gear of planetary gearbox is stationary and the sun gear is the input of the planetary gearbox system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001428_s11831-014-9106-z-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001428_s11831-014-9106-z-Figure11-1.png",
        "caption": "Fig. 11 Lower leg force diagram for muscle-less model",
        "texts": [
            " In such a case, the equilibrium of the lower leg can be solved for the unknown knee forces, Fkx2 (normal component) and Fky2 (tangential component), and moment, Mk . Fkx2 = \u2212m2gsin(\u03b12) + Fsin (\u03b12) (17) Fky2 = m2gcos(\u03b12) \u2212 Fcos (\u03b12) (18) Mk = m2gL2Cm2cos (\u03b12) \u2212 F L2cos (\u03b12) (19) Equilibrium equations 17 and 18 are derived in the lower leg\u2019s local coordinate system, while moment equilibrium equation 19 is derived in the global coordinate system. This arrangement makes derivation straightforward. Figure 11 illustrates lower leg segment model used in the derivation of Eqs. 17\u2013 19. 22.2 Model B: Muscle Actuation of the Knee Modeling a joint with muscles requires the separate derivation of knee forces. In this model, the knee joint is regarded as a pin joint, thus no internal knee moment exists. However, to restrain the lower leg from swinging the role of knee moment is taken by the muscle forces. Depending on the moment sign in the muscleless model, either the knee extensors or flexors are active. Therefore, to solve for the unknown muscle forces, the moment computed earlier can be used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003384_iscid.2016.1087-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003384_iscid.2016.1087-Figure1-1.png",
        "caption": "Figure 1. Two-mass system transmission structure",
        "texts": [
            " This kind of periodic motion is called mechanical resonance, whose emergence not only causes instability of servo motor and deteriorates control performance, but also jeopardizes the transmission mechanism under severe circumstance, which eventually reduces the motor life and endangers industrial production. II. TWO-MASS MATHEMATICAL MODEL In the elastic servo system, the mechanical stiffness of its transmission shaft is limited. The drive motor and load can be equivalent into two large inertia objects, so is transmission part as a spring with viscous coefficient. According to analysis in [4], we can simplify the system into two-mass mathematical model so as to explore the theoretical cause of mechanical resonance. 2473-3547/16 $31.00 \u00a9 2016 IEEE DOI 10.1109/ISCID.2016.86 349 As is shown in Fig. 1, 1J and 2J represent inertia moment of the motor and the load, respectively. eT , LT and WT refer to the electromagnetic torque, the load torque and the shaft torque, respectively. 1B , 2B and \u03c9B are damping coefficient of the motor, the load and the shaft, respectively. 1\u03b8 and 2\u03b8 are angular displacement of the motor and the load, respectively. K is the stiffness of the shaft. The kinetic equation of two-mass system can be established based on the model in (1). 1 1 e W 1 1 2 2 W L 2 2 W \u03c9 1 2 1 2 J \u03b1 T T B \u03c9 J \u03b1 T T B \u03c9 T B \u03c9 \u03c9 \u03b8 \u03b8K where 1\u03c9 and 2\u03c9 denote angular speed of motor and load, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.25-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.25-1.png",
        "caption": "Fig. 35.25 Oil bath lubrication.",
        "texts": [
            " Bearings and applied technology 543 Oil lubrication is generally used where high speeds or operating temperatures prohibit the use of grease, when it is necessary to transfer frictional heat or other applied heat away from the bearing, or when the adjacent machine parts, for example gears, are oil lubricated. Oil bath lubrication is only suitable for slow speeds. The oil is picked up by rotating bearing elements and after circulating through the bearing drains back to the oil bath. When the bearing is stationary, the oil should be at a level slightly below the center of the lowest ball or roller. An application is shown in Fig. 35.25. At high speeds it is important that sufficient oil reaches the bearing to dissipate the heat generated by friction and oil jets provide an effective method (Fig. 35.26). The illustrations in this section (Fig. 35.10e35.26) are reproduced by kind permission of SKF (UK) Limited e www.skf.co.uk. 544 Manual of Engineering Drawing When bearings fail, they can bring equipment to an unscheduled halt. Every hour of down time due to premature bearings failure can result in costly lost production in a capital intensive industry"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000484_s00574-015-0103-0-Figure3.27-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000484_s00574-015-0103-0-Figure3.27-1.png",
        "caption": "Figure 3.27: The first picture shows the graph of the funcition \u03b8\u0303(C) for a = \u22121. The first point in the graph (c1(\u22121), b1(\u22121)) has been highlighted as well as the point on the graph with second entry equal to \u03c0 4 . The profile curve of immersed surface represented by the point (0.164021, \u03c04 ) is shown in the last picture. The picture in the middle shows the graph of the function b1(a) which represents the limit of the function \u03b8\u0303(C) when C goes to the first value in its domain. The point on this graph when a = \u22121 has been highlighted.",
        "texts": [],
        "surrounding_texts": [
            "In this subsection we will find some embedded examples and we will display their profile curves. Many of these examples have appeared elsewhere, e.g. [9], [12]. As pointed out before, when the cylindrical drop is not exceptional, its profile curve is a union of rotations of fundamental pieces that ends up being invariant under the group G define in 3.6. It is not difficult to see that a necessary condition for the cylindrical drop to be embedded is that \u03b8\u0303 = 2\u03c0/m for some integer m. We will show that this condition is not sufficient. In order to obtain this potentially embedded example we need to understand the function \u03b8\u0303(C, a, x1, x2). We will apply the Intermediate Value Theorem in order to solve the equation \u03b8\u0303 = 2\u03c0/m . We know that for any a, there is a first (or last) value of C, which we denote by C0(a, x1, x2), for which the function \u03b8\u0303 is define. We will compute the limit of \u03b8\u0303 when C goes to C0 using Lemma 3.6. Before this, we will show images of some of the embedded examples. These graphs were generated by the software Mathematica 8, by solving the system of ordinary differential equations (see equation (2.4)) x \u2032(t) = cos(\u03b8(\u03b8)), y \u2032(t) = sin(\u03b8(\u03b8)), \u03b8 \u2032(t) = \u2212 0 + a 2 (x(t)2 + y(t)2). Proposition 3.13. Let r1, r2 and r3 be the functions given in Definition 2 and let C1, C2 and C3 be the functions defined in Lemma 3.11. For i = 1, 2, define the functions bi (a) = (\u03c0(4Ci + ri (ari \u2212 4))) ri \u221a 16 + 8a(Ci \u2212 3ri )+ 6a2r2 i , and let x1 and x2 be the first two roots of the polynomial q(r, a,C). Bull Braz Math Soc, Vol. 46, N. 4, 2015 a) If a < 0, then limC\u2192C1 (a)+ \u03b8\u0303(C, a, x1, x2) = b1(a). b) If a > 0 then limC\u2192C1 (a)\u2212 \u03b8\u0303((C, a, x1, x2) = b1(a). c) If a > 0 then limC\u2192C2 (a)+ \u03b8\u0303((C, a, x1, x2) = b2(a). Proof. Since \u03b8\u0303 = \u222b x2 x1 4C+r(ar\u22124) r \u221a q , when C approaches the limit value the two roots approaches ri (i = 1 or 2) which are roots of q with multiplicity 2. Therefore Lemma 3.6 applies and the proposition follows. Notice that the value A in Lemma 3.6 is given by A = \u22121 2 q \u2032\u2032(ri ) = 16 + 8a(Ci \u2212 3ri )+ 6a2r2 i . Consider the graph of C \u2192 \u03b8\u0303 for some particular value of a in the cases considered in Proposition 3.12. When a is between 0 and 8 27 and C is between C2(a) and C3(a), there are two surfaces associated with these values of a and C due to the fact that the curve G = C has two connected components. These surfaces are those described in case (ii) of Proposition 3.12. A natural question to ask is if it is possible to find values of a and C such that both surfaces are properly immersed, that is, can we find a and C such that \u03b8\u0303(C, a, x1, x2)/2\u03c0 and \u03b8\u0303(C, a, x3, x4)/2\u03c0 are rational numbers. Here x1 < x2 < x3 < x4 are the four roots of q(r,C, a). Bull Braz Math Soc, Vol. 46, N. 4, 2015 Bull Braz Math Soc, Vol. 46, N. 4, 2015 As a consequence of the following proposition we have that either both surfaces are properly immersed or both surfaces are dense. Proposition 3.14. Let 0 = 1, let a \u2208 (0, 8/27), let C \u2208 (C2(a),C3(a)) and let x1 < x2 < x3 < x4 be the four roots of q(r, a,C) = \u221216C2 + 64r + 32C r \u2212 16 r2 \u2212 8aC r2 + 8a r3 \u2212 a2 r4. Then if C < 0, there holds \u03b8\u0303(C, a, x3, x4) = \u222b x4 x3 (4C + ar2 \u2212 4 0r) r \u221a q(r, a,C) dr = \u222b x2 x1 (4C + ar2 \u2212 4 0r) r \u221a q(r, a,C) dr + 2\u03c0 = \u03b8\u0303(C, a, x1, x2)+ 2\u03c0 . If C > 0 then there holds \u03b8\u0303(C, a, x3, x4) = \u222b x4 x3 (4C + ar2 \u2212 4 0r) r \u221a q(r, a,C) dr = \u222b x2 x1 (4C + ar2 \u2212 4 0r) r \u221a q(r, a,C) dr = \u03b8\u0303(C, a, x1, x2) . Bull Braz Math Soc, Vol. 46, N. 4, 2015 Proof. Let U be the region shown in Figure 3.29 bounded by the curves \u03b31, \u03b32, \u03b33 and \u03b34. The curves \u03b3i will carry the orientations induced by the fact that they are components of \u2202U . Define f (z) := i(4C + az2 \u2212 4z) za(z \u2212 x2) \u221a T1(z)(z \u2212 x4) \u221a T2(z) , where \u221a is the principal branch of the square root, T1(z) = z\u2212x1 z\u2212x2 and T2(z) = z\u2212x3 z\u2212x4 . Bull Braz Math Soc, Vol. 46, N. 4, 2015 When the radius of \u03b31, and the radii of the circular arcs of \u03b32 and \u03b33 tend to zero, we have that \u222b \u03b31 f (z) dz goes to \u22122\u03c0 C |C| , \u222b \u03b32 f (z) dz goes to 2 \u03b8\u0303(a,C, x1, x2) and \u222b \u03b33 f (z) dz goes to \u22122 \u03b8\u0303(a,C, x3, x4). When the radius of \u03b34 goes to \u221e, we have that \u222b \u03b34 f (z) dz goes to \u22122\u03c0 . The result follows because the integral\u222b \u03b31+\u03b32+\u03b33+\u03b34 f (z)dz vanishes since f is analytic in U . Bull Braz Math Soc, Vol. 46, N. 4, 2015"
        ]
    },
    {
        "image_filename": "designv11_64_0000651_isie.2015.7281516-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000651_isie.2015.7281516-Figure3-1.png",
        "caption": "Fig. 3. Bearings with holes: (a) single 2 mm inner race defect and (b) single 4 mm outer race defect.",
        "texts": [
            " A 4 kW dc machine operating as a generator was used to impose different load levels to the motor under test: no load, half load and full load. One of induction motors was brand new and was tested with its original bearings, being the results obtained used as a reference for a motor without defects (healthy). The second induction motor had its bearings replaced by others with holes produced in its internal and external rings for the emulation of bearing faults. The produced holes have diameters of 2 and 4 mm for simulating different degrees of damage to the bearings. Fig. 3 shows two examples of bearings with holes introduced to simulate defects in both the inner and the outer ring. The bearings used in the tested motors were type 6205 (non-drive end) and 6206 (drive end). The motor vibration signals were measured using Vibrasens triaxial vibration sensors, model 131, with a sensitivity of 1000 mV/g. The vibration sensors were installed near the drive and non-drive ends. The instrumentation is complemented with a NI CompactRIO 9074 data acquisition system and the signals processing using Labview software"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000542_1.4928573-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000542_1.4928573-Figure2-1.png",
        "caption": "FIG. 2. A cross section (drawn for generality as rectangular) with its p1p2t frame (the small squares at the origin indicate that the tangent t to the space curve (not shown) is perpendicular to the cross section). The right-handed screw rotation arrows show what happens when the cross section translates infinitesimally along the axial space curve. When the cross section is rotated about each of the vectors pi, the cross section at s+ds is tilted with respect to the one shown at s and the axial space curve must bend. When the cross section is rotated about t, the translated cross section is twisted with respect to the one shown.",
        "texts": [
            " Moreover, if the rod is sufficiently thin, we can neglect the deviation of the cross section from planarity (because the tangent plane spanned by the frame may be regarded as touching the cross section in a small area near the origin). The transverse cross sections of a circular rod do remain planar in a twisting deformation.2 In some treatments, a cross section of a thin rod is simply defined by the p1p2 frame but that seems unnecessarily abstract.10 The deviation from planarity manifests itself energetically in the elastic twisting constant C, which in our application is a measured quantity. We now arrive at an expression for the components of the rotation vector \u2126 first by an intuitive geometric argument based on Figure 2 and then by a formal calculation. In Figure 2, we show a cross section of the rod (taken for generality as rectangular), along with its embedded frame p1p2t. The cross section is supposed to reside at arc length s along the axis. The cross section at s + ds is obtained by rotating the cross section shown in the figure (i.e., tilting and twisting it while translating from s to s + ds), in a manner consistent with the shape of the rod. Referring to Figure 2, let us first view p1 as an axis of rotation for the two-dimensional frame p2t with the usual sense of a right-handed screw, as indicated in the figure. It should be clear (possibly with some effort at internal visualization) that this rotation results in a tilt of the cross section away from the direction of p2. The tilt means that the rod bends. The corresponding component of bending is then \u2212\u03ba2, the negative of the bending of the axis in the direction p2. (The components of bending \u03ba1 and \u03ba2 will be presented more systematically in the formal calculation to follow.) If\u2126 is resolved in the frame p1p2t, \u2126(s) = \u21261(s)p1(s) +\u21262(s)p2(s) +\u2126t(s)t(s), (1) we then have\u21261 = \u2212\u03ba2. Similarly, if we view p2 in Figure 2 as the axis of rotation of the two-dimensional frame tp1, the result is a bend toward p1, so that \u21262 = \u03ba1. At this point we would like to review the reason why these components of \u2126, when they describe deformation of the rod away from its undeformed state, are elastic deformations and thus generate elastic bending energy. Longitudinal fibers in the rod passing through the cross section in the portion of it away from the bend are stretched by the bending deformations just described. Those passing through the portion on the inside of the bend are contracted"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001650_9781782421955.999-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001650_9781782421955.999-Figure2-1.png",
        "caption": "Figure 2: principle of the damping device",
        "texts": [
            " An example of a Campbell Diagram of a power gear is given in figure 1. In some cases, resonances may induce unacceptable noises, vibration or stress. In these cases the durability of the gears may be significantly affected. A damping device has been developed for gears which can reduce significantly the vibration level at resonance. It has been used in Turbomeca\u2019s gears in the 90\u2019s and before. The technology retained is a ring which is introduced into a groove under the rim of the gears. The principle is presented in figure 2. 2.1 Basic principle A damping ring is introduced in a groove under the rim of the gear with a light prestress. In rotation, due to centrifugal loads, the ring is stuck up to the rim. On a resonance, due to the deformed shape of the rim, a sliding between the rim and the ring occurs. This sliding introduces damping in the system. 2.2 Existing bibliography and article original work In ref (3), an expression of the dissipated power in case of an axisymmetric structure with N nodal diameter modes is given (example of labyrinth seal)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure9.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure9.7-1.png",
        "caption": "Fig. 9.7 Transient analysis of a rod structure: a schematic sketch of the problem and discretization; b force-time relationship",
        "texts": [
            "53) to solve for u1: u1 = 1 M [ t2Fx (0) \u2212 t2 ( K \u2212 2M t2 ) u0 \u2212 Mu\u22121] = 0.0833\u0304. 5 Use Eq. (9.40) to solve for u\u03081: u\u03081 = 1 M (Fx(t1) \u2212 K u1) = 49.722\u0304. 6 Use Eq. (9.45) to solve for u\u03071: u\u03071 = u2\u2212u0 2 t = 2.909722\u0304. 7 Repeat steps 5\u20137 to obtain displacement, acceleration, and velocity for all other time steps. 9.2 Example: Transient analysis of a rod structure\u2014discretization via two finite elements, consistent and lumped mass approach Consider a cantilevered rodwhich is discretizedwith two finite elements, see Fig. 9.7. The geometry and the material behavior are described by A = 650, L = 2540, E = 7.8 \u00d7 10\u22129, and E = 210000. A constant load of Fx = 4450 is applied on the right-hand end of the rod. Calculate the nodal displacements, velocities and accelerations for 0 \u2264 t \u2264 0.001 and t = 0.00025 based on the solution scheme from p. 364. 9.2 Solution (a) Consistent mass approach for M Element I: AL 6\ufe38\ufe37\ufe37\ufe38 m 6 [ 2 1 1 2 ] [ u\u03081x u\u03082x ] + E A L [ 1 \u22121 \u22121 1 ] [ u1x u2x ] = [\u2212R1 0 ] . (9.70) Element II: AL 6\ufe38\ufe37\ufe37\ufe38 m 6 [ 2 1 1 2 ] [ u\u03082x u\u03083x ] + E A L [ 1 \u22121 \u22121 1 ] [ u2x u3x ] = [ 0 Fx (t) ] "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002935_gt2016-56508-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002935_gt2016-56508-Figure14-1.png",
        "caption": "Figure 14: Free body diagram of a beam model",
        "texts": [
            " The imbalance responses at both bearing locations are shown in Figure 15 and Figure 16 along with other simulation results. The analysis reveals rigid mode critical speed at 20krpm at the Fwd AFB and at 40krpm at the Aft AFB. The difference in critical speeds observed at two bearing locations is due to the fact that the majority of the rotor weight is supported by the Fwd AFB. Steady state vibration magnitude is about 2\u03bcm at Aft AFB and 1\u03bcm at Fwd AFB. Linear Flexible Shaft Model: The rotor components can be characterized as stiffness elements or added inertia elements as shown in Figure 13 and Figure 14. Particularly, the shaft beam elements are stiffness elements; and those of the thrust runner, overhung mass and impulse turbine are modeled as added inertia elements. The steady state solution of the imbalance response is also presented in Figure 15 and Figure 16. At the Fwd bearing location, the magnitude of rotor vibration is much smaller than the prediction from the previous rigid rotor model. The analysis also shows a peak at 190krpm and predicts excessive vibration when the rotor speed is above 200krpm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.6-1.png",
        "caption": "Fig. 3.6 Example for spherical kinematic chains: transmission linkage with CARDAN joint",
        "texts": [
            " Kinematic chains can be grouped into three distinct motion categories: Planar kinematic chains In a planar kinematic chain all body points move inside or parallel to a reference motion plane (Fig. 3.5). Because of this, the motion of each body in the system has one rotational and two translational motion components. Relative motions between joints must be either translational displacements, that are parallel to the reference motion plane, or rotations, that are normal to the motion plane. Spherical kinematic chains In a spherical kinematic chain, all body points move on concentric spherical surfaces around a fixed point O in the center (Fig. 3.6). The bodies have three rotational motion components and no translational components. The relative motions of the bodies can only rotate around the axes going through the fixed point. Spatial kinematic chains In spatial kinematic chains, the motion of the bodies can be described as general spatial. In general there are three translational and three rotational motion components. The relative motion of the bodies in the joints\u2014depending on the joint\u2014is as well arbitrary. This is illustrated with the complex example of a five-point wheel suspension in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure3.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure3.8-1.png",
        "caption": "Fig. 3.8 In equilibrium, the applied force is balanced by the net force \u03b4FN between the atoms as a result of their increased separation",
        "texts": [
            " More generally, even for a solid (not only for metals) in the presence of many interacting atoms, we can still identify a general potential energy and force curves per atom similar to the type shown in Fig. 3.7a. We can examine these curves for a better understanding of the properties of the solid, such as the thermal expansion coefficient and modulus of elasticity. When a solid is under an external stretching force, the applied stress causes two neighboring atoms along the direction of force to be further separated. Their displacement \u03b4r (\u00bcr ro) results in a net attractive force \u03b4FN between two neighboring atoms as indicated in Fig. 3.8, where FN is the net interatomic force. \u03b4FN attempts to restore the separation to equilibrium and is balanced by a portion of the applied force acting on these atoms with an effective area. Such area for each atom would be roughly ro 2. For an ideal pure metallic body without defects and impurities, if we consider \u03b4FN/ro 2 as the average \u201catomic\u201d stress and \u03b4r/ro as the average \u201catomic\u201d strain, we may relate these quantities with the bulk elastic modulus E of a material body (defined in Eq. 2.4) as \u03b4FN/ro 2 \u00bc E\u03b4r/ro or E 1 ro dFN dr r\u00bcro : \u00f03:1\u00de Clearly, E depends on the gradient of FN versus r curve at ro in the plot indicated in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002196_1468087414556134-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002196_1468087414556134-Figure13-1.png",
        "caption": "Figure 13. Refitted single-cylinder air compressor.",
        "texts": [
            "0 order frequency is influenced by the friction force of piston rings, which can be used to judge scuffing failure. To validate the conclusion obtained from the simulation results, an experiment of scuffing failure is needed to be implemented. However, scuffing failure is a serious fault for diesel engine, and it is difficult to simulate in the actual engine. Furthermore, the scuffing failure once appears in the actual engine, which will destroy the engine and threaten the safety of the people. So, in this study, the experiment is implemented in a refitted single-cylinder air compressor as shown in Figure 13. The air compressor is also a reciprocating machine, which consists of piston assembly, cylinder liner, connecting rod and crankshaft. The working process of air compressor contains intake, compression and exhaust except expansion. So, the air compressor has the same movement mechanism as diesel engine and experiences similar working process, which is suitable for simulating scuffing failure instead of diesel engine. As shown in Figure 13, the air compressor is driven by an electromotor, and a gear and a magnetoelectric sensor are installed at the end of electromotor to measure torsional vibration. The parameters of air compressor are listed in Table 8. The schematic diagram of measurement system of torsional vibration is shown in Figure 14. A gear is installed at the free end of the electromotor, which rotates with the shaft synchronously, and a magnetoelectric transducer is used to acquire the interval of each tooth of the gear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003101_icelmach.2016.7732702-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003101_icelmach.2016.7732702-Figure2-1.png",
        "caption": "Fig. 2. Fem flux lines on load with six-step rectangular currents.",
        "texts": [
            " Non-intrinsic UMF The r- and \u03b8-components of magnetic pressures (in function of r\u03b8 and \u03b8 ) at gR are defined as [10] 22 0 1 2 g g r rI Ir R r RP B B\u03b8\u03bc = == \u22c5 \u2212 (43) 0 1 g g rI Ir R r RP B B\u03b8 \u03b8\u03bc = == \u22c5 \u22c5 (44) The x- and y-components of non-intrinsic UMF at gR is calculated as [3] and [7] ( ) ( ) 2 22 0 0 cos 2 2 sin gg gg I rI r Rr Rg u x I rI r Rr R B BR L F d B B \u03c0 \u03b8 \u03b8 \u03b8 \u03b8 \u03bc \u03b8 == == \u2212 \u22c5\u22c5 = \u22c5 \u22c5 + \u22c5 \u22c5 \u22c5 (45) ( ) ( ) 2 22 0 0 sin 2 2 cos gg gg I rI r Rr Rg u y I rI r Rr R B BR L F d B B \u03c0 \u03b8 \u03b8 \u03b8 \u03b8 \u03bc \u03b8 == == \u2212 \u22c5\u22c5 = \u22c5 \u22c5 \u2212 \u22c5 \u22c5 \u22c5 (46) V. ANALYTICAL RESULTS AND FEM VALIDATION The main dimensions and parameters of the fractionalslot STPM machine (6-slots/4-poles) with the buried PMs and a single layer winding (i.e., non-overlapping alternate teeth wound winding) are given in Table I. Then, analytical results are verified by 2-D Fem [16]. Fig. 2 shows that the flux lines are not equilibrate in each part of the machine in the case of rectangular drive currents. This can be also confirmed in Fig. 3, where radial and tangential flux densities have a periodicity equal to 2 . Figs 4 ~ 6 show the bach-EMF, electromagnetic torque (for both drive currents) and cogging torque curves. We can see in the spectrum of back-EMF [see Fig. 4.b] that the odd harmonics orders 1, 5, 7 and 10 are predominates. For cogging torque spectrum, even harmonics orders 2, 4, 8, and 10 are significant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.39-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.39-1.png",
        "caption": "Fig. 3.39 Different modelling approaches for a bending problem: a Problem sketch; b 1D beam elements; c 2D plane elasticity elements; d 3D solid elements",
        "texts": [
            " \u2022 State the required (a) geometrical parameters and (b) material parameters to define an Euler\u2013Bernoulli beam element. \u2022 Sketch the interpolation functionsN1(\u03be),N2(\u03be), andN3(\u03be)of anEuler\u2013Bernoulli beam element. \u2022 Sketch the shape functions N1(\u03be) and N2(\u03be) of an Euler\u2013Bernoulli beam element. \u2022 Explain in words the difference between an Euler\u2013Bernoulli beam element and a generalized beam element in regards to the nodal unknowns. \u2022 State the DOF per node for a generalized beam element in a plane (2D) problem. \u2022 State the DOF per node for a generalized beam element in a 3D problem. \u2022 Figure3.39a shows schematically a cantilevered Euler\u2013Bernoulli beam. In a finite element approach, such a beam can be modeled based on one-dimensional beamelements (Fig. 3.39b), two-dimensional plane elasticity elements (Fig. 3.39c), or three-dimensional solid elements (Fig. 3.39d). State for each approach one advantage. \u2022 Figure3.40 shows a plane frame structure which should be modeled with three generalized beam (I, II, III) elements. State the size of the stiffness matrix of the non-reduced system of equations, i.e. 172 3 Euler\u2013Bernoulli Beams and Frames without consideration of the boundary conditions. What is the size of the stiffness matrix of the reduced system of equations, i.e. under consideration of the boundary conditions? 3.11 Cantilevered beam with a distributed load: analytical solution Calculate the analytical solution for the deflection uz(x) and rotation \u03d5y(x) of the cantilevered beam shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000075_gt2012-68354-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000075_gt2012-68354-Figure5-1.png",
        "caption": "Figure 5. The inside of the bearing chamber",
        "texts": [
            " 2) pressure, temperature and mass flow measurements of the sealing air upstream of each seal individually. 3) temperature and pressure measurements of the air/oil mixture in the bearing chamber. 4) oil flow data and bearing control. Downstream of the brush seal, a wind back seal was attached to the rotor as a part of a combined sealing package for bearing chambers. The function of the windback seal [1] was to inhibit oil migration from the bearing chamber and therefore support the sealing function of the brush seal. Figure 5 depicts the inside of the bearing chamber with the locations A and B where the transient temperature recording takes place through the pyrometers. Part of the pyrometer calibration process was also to darken the areas facing the pyrometers using special paint. In order to measure the rotor surface temperatures at the contact zone, standard thermocouple instrumentation could be used. Nevertheless, the respective effort would be remarkably high, since either a telemetry or a slip ring system with all respective disadvantages would be necessary"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003908_1.1698360-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003908_1.1698360-Figure2-1.png",
        "caption": "FIG. 2. Elastic curve of a single metal beam.",
        "texts": [
            " On the other hand, the beams give a positive restoring moment because of their natural rigidity. Therefore it seems evident that for a suitable choice of all proportions, especially of P\" the total restoring moment of the bearing can be made zero or even negative. The following discussion gives a short quantita tive representation of the relations wherefrom all consequent requirements for the design of such bearings can be derived. For this purpose the elastic curve of a single metal beam is represented in Fig. 2. The tangents to the elastic curve at the ends I and II correspond to both radii crossing in the center of motion O. In the case in question the circle B has been turned through the angle cp out of its original position. Also, since in the practical case of application the angle cp will always be very small (less than 1\u00b0), the curvature of the beam can be approximated for the following calculation with sufficient accuracy by the second differential quotient of the elastic curve. Then the following differential equation of the beam under axial tensile stress results (d2y/dx2) - (P\"j I\u00b7E)\u00b7y = -(Py/I\u00b7E)\u00b7x+(Mr/I\u00b7E), (1) where I=moment of inertia of the cross section of the beam with respect to the neutral axis",
            "15(E/S)D\u00b7\u00b7\u00b7b\u00b7 cpo (11) Equation (11) provides for precalculation of the torque due to a small deviation of the tensile force p. from its prescribed value P'o. Practically, these changes can never be avoided so that such a bearing always has a certain small torque. Especially, any changes in temperature can affect such a bearing. Moreover the tensile force P. decreases with in creasing angle cpo This is due to the fact, that the necessary beam length becomes smaller when the point II moves upon the circle with R2 around the center 0 (see Fig. 2). This change in length of the elastic curve between the points I and II results in a decrease of P., the consequence of which, accord ing to Fig. 3, is a positive restoring torque. Ob viouslysuch a bearing has only a negligible re storing torque in the immediate surroundings of VOLUME 20, APRIL, 1949 rp=O, which, however, increases with cp to positive values as P. varies. These disadvantages can be nearly compensated for if a second bearing under compression stress is used. The principal arrangement of such a bearing exactly equals the one shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000816_s10409-015-0016-6-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000816_s10409-015-0016-6-Figure13-1.png",
        "caption": "Fig. 13 Contours of the dimensionless axial velocity V\u03c2 on various planes: a \u03b7 = 0; b \u03be = 0; c \u03c2 = 0.50; d \u03c2 = 1.0",
        "texts": [
            " These properties reflect the physical essence pertinent to the problem and are graphically shown in Fig. 11. As expected, the contours on the horizontal planes (Fig. 11) are not concentric circles. The dimensionless velocity component V\u03be is displayed in Fig. 12. We can infer from the first expression in Eq. (42) that vx(x, y, z) = \u2212vx(\u2212x, y, z), vx(x, y, z) = \u2212vx(x, y,\u2212z), vx(x, y, z) = vx(x,\u2212y, z), vx(x, y, z) = vx(\u2212x,\u2212y, z). These interesting relations are reflected in Fig. 12. It is seen that the velocity changes significantly in the neighborhood of the plate. Figure 13 plots the distributions of the dimensionless velocity component V\u03c2. The symmetry with respect to the xOz and yOz planes are observed. However, no symmetric behaviors can be found with respect to plane z = 0 as indicated by the expressions of vz in Eq. (42). Apparently, the pressure and velocity components vary significantly in the vicinity of the elliptic plate. The fluid flow at the points with a distance 4a from the center of the plate becomes even. Hence, the corresponding contours become circular, as shown by Figs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.3-1.png",
        "caption": "FIGURE 8.3",
        "texts": [
            " When the system is conservative, Q \u00bc 0. For a nonconservative system (for example, if friction is present), Q \u00bc F, where F is the vector of generalized forces. For the simple pendulum example, q \u00bc [q], L \u00bc T U \u00bc 1 2m\u2018 2 _q 2 mg\u2018\u00f01 cos q\u00de, and Q \u00bc 0. Thus, using Eq. 8.8, we have d dt m\u20182 _q \u00f0 mg\u2018 sin q\u00de \u00bc m\u20182\u20acq\u00fe mg\u2018 sin q \u00bc 0 (8.9) Therefore, \u20acq\u00fe g \u2018 sin q \u00bc 0, which is the same as Eq. 8.2 (and Eq. 8.3 if q z 0). Now consider a particle sliding along a slope defined by a parametric curve x(u), shown in Figure 8.3(a), with no friction. Note that the curve can be planar or spatial. The position, velocity, and acceleration of the particle can be obtained by solving the equation of motion derived from the Lagrange equation. The kinetic energy and potential energy are, respectively, T \u00bc 1 2 m _x2 \u00bc 1 2 m\u00f0x; u _u\u00de2 and U \u00bc mgx2(u), where m is the particle mass and x,u \u00bc vx/vu. The Lagrangian is L \u00bc 1 2 m\u00f0x; u _u\u00de2 mgx2\u00f0u\u00de: Using Eq. 8.8, the equation of motion can be derived as x;2u\u20acu\u00fe 2x;ux;uu _u 2 \u00fe gx2;u \u00bc 0 (8",
            " Note that the equation of motion shown in Eq. 8.10 can be used in applications such as rollercoasters, where position, velocity, and acceleration of riders modeled as concentrated mass are calculated. Equation 8.10 can be extended to support particle motion on a spatial parametric surface x(u,w) for applications such as waterslides and bobsleds. More on this topic is given in Section 8.6. Note that, in general, Eq. 8.10 must be solved using a numerical method. Only very simple casesdfor example if x(u) is a straight line as shown in Figure 8.3(b)dcan be solved analytically. The following example illustrates the details. EXAMPLE 8.2 A particle slides along a straight line, as shown in Figure 8.3(b), from p0 to p1, where p0\u00bc [0,1] and p1\u00bc [1,0]. Calculate the position, velocity, and acceleration of the particle and the time required for it to reach point p1. Assume the initial velocity to be zero. Solution The parametric equation of a straight line is x\u00f0u\u00de \u00bc p0\u00f01 u\u00de \u00fe p1u \u00bc \u00bdu; 1 u (8.11) Object sliding along a curve x (u): (a) general curve and (b) straight line. EXAMPLE 8.2econt\u2019d Also, x;u\u00f0u\u00de \u00bc \u00bd1; 1 ; x;uu\u00f0u\u00de \u00bc 0; \u00f0x;u\u00f0u\u00de\u00de2 \u00bc x;u\u00f0u\u00dex;u\u00f0u\u00deT \u00bc 2; and x2;u \u00bc 1 Therefore, the equation of motion becomes \u20acu 1 2 g \u00bc 0 (8",
            " It is hoped that this chapter has provided the reader with a good understanding of how the motion analysis method works, how to create motion analysis models, and how to choose the right software tool for the problem at hand. The chapter also presented case studies involving design aspects of motion analysis, including a Formula SAE racecar and an HMMWV suspension. These two cases should provide a general idea of the applications that lend themselves to simulation for motion analysis and design in general. QUESTIONS AND EXERCISES 8.1. Derive equations of motion for the particle sliding along a straight line shown in Figure 8.3(b) using Newton\u2019s method. Compare your results with those obtained in Example 8.2. 8.2. Repeat Exercise 8.1; that is, derive equations of motion for the particle shown in Figure 8.3(b) using Newton\u2019s method as well as Lagrange\u2019s equations, assuming that friction coefficient m is nonzero. 8.3. Equation 8.10 was derived with an assumption of no friction. If friction is present, derive the generalized forces Q for the particle sliding along the curve x(u). 8.4. Repeat Example 8.3 using an initial angular velocity that is greater than umin \u00bc ffiffiffi g r0 q , for example, u0 \u00bc ffiffiffiffiffi 8g 3r0 q . 8.5. Derive the acceleration equations for the slider-crank mechanism by taking derivatives of Eqs 8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.14-1.png",
        "caption": "Fig. 35.14 Single and double row cylindrical roller bearings.",
        "texts": [],
        "surrounding_texts": [
            "The chief characteristic of needle roller bearings is that they incorporate cylindrical rollers with a small diameter/length ratio. Because of their low-sectional height these bearings are particularly suitable for applications where radial space is limited. Needle roller bearings have a high load carrying capacity in relation to their sectional height."
        ]
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure5-1.png",
        "caption": "Figure 5 A schematic of tensile test specimen configurations (see online version for colours)",
        "texts": [
            " Carefully calibrations of load cell and the biaxial extensometer were carried out before experiments. Tensile test specimens were fabricated in accordance with the ASTM D638 (2003) standard test method for tensile properties of plastics. The gauge dimension is 50.80 mm (2.00 inch) in length and 12.70 mm (0.50 inch) in width with 7.11 mm (0.28 inch) in thickness. The specimens were fabricated in three configurations as those for the torsion test specimens in order to obtain mechanical properties in three directions as shown in Figure 5. Each sample had five replicas for tensile tests. The test matrix for tensile test and the results (to be discussed later) are shown in Table 2. It should be noted that the moduli of TN4 and TN5 specimens were expected to be approximately equal. It was also expected that applying 45\u00b0 rotation of the obtained compliance matrix about the z-axis using transformation laws would obtain the values of the moduli of TN1 and TN2. In the experiments, it was observed that the material properties (e.g. shear and Young\u2019s moduli) of the specimens vary with their building configurations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000990_amm.555.192-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000990_amm.555.192-Figure1-1.png",
        "caption": "Fig. 1. Illustration of programmed and corrective speed components",
        "texts": [],
        "surrounding_texts": [
            "Calculation of trajectories for both robots is done according the same algorithm. The algorithm described in the text is restricted to Cartesian positions. The version of the algorithm which includes orientation is presented in [4]. It is assumed that position PA of robot A 1 TCP is calculated as the sum of two motion components: \u2022 the programmed motion \u2013 this component represents the motion with constant programmed velocity along a vector from current robot A TCP position to the set final position , \u2022 the corrective motion \u2013 this component represents the motion which aims at minimizing the trajectory error. The velocity of this motion is set along the direction which connects current robot A TCP position with current robot B TCP position . An illustration of the programmed and corrective speeds for 2D case and two different positions As a measure of trajectory error E the change of distance between robots TCPs or force interaction may be used. Block diagram which presents calculation of programmed and corrective speeds is shown in Fig. 2. 1 Motion parameters like TCP positions and speeds are functions of time. For the reason of notation simplicity it is not marked explicitly in the text and formulas. TCP positions of robot A and B (initial, current and final) are defined in robot A and B coordinate system respectively. Applied Mechanics and Materials Vol. 555 193 Grey rectangles represent calculation blocks, while dashed lines data flow. The vector of programmed speed is calculated according to Eq. 1. = (1) Trajectory error defined as change of distance between robots TCPs is calculated according Eq. 2. = ( \u2212 + ) \u2212 | | ( \u2212 + ) (2) where: , \u2013 initial positions of TCP of robot A and B, \u2013 vector defining position of robot B coordinate system in robot A coordinate system (the same orientation of both coordinate systems is assumed). The corrective speed is calculated on the basis of estimated trajectory error using proportional and integral rule presented by Eq. 3. = ( + ) (3) where: and are gain and integral constant of the corrector respectively. Modification of programmed motion component Instead of a simple programmed motion definition with speed direction from current to the final position (in general this direction is not constant in time because of corrective motion component) programmed trajectory may be set by the user to influence the generated trajectory. In such a case is a predefined curve which starts in the initial position and ends in the final position . The tangent to the curve defines direction of temporary programmed speed . The current position of robot A TCP does not coincide with the programed position .Thus the programmed speed which results from the programmed curve geometry cannot be used directly as a programmed speed component to calculate the speed along the generated trajectory. It needs to be modified. For calculation of modified programmed speed component two methods are proposed. In Fig. 3 a modification of programmed speed component based on geometric approach is presented."
        ]
    },
    {
        "image_filename": "designv11_64_0003308_ichve.2016.7800628-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003308_ichve.2016.7800628-Figure1-1.png",
        "caption": "Figure 1 Simulation model of the power tower",
        "texts": [
            " A 220kV straight line tower was chosen as the simulation model. Firstly, a simulation model consists of the transmission line and UAV is built in the finite element analysis software (Ansoft Maxwell). In order to make the simulation more typical, a 220kV power tower (type: 2E1SZ1) which is shown in Figre.1 is selected as the simulation object. 978-1-5090-0496-6/16/$31.00 \u00a92016 IEEE Based on the manufacturing standard [7], the main structure parameter of the 2E1SZ1 tower and composite insulator is shown in Figure.1 and Table.1 respectively. Actually, the whole tower is zero potential and its size is much larger than that of the UAV and transmission line. The model of UVA and insulator are shown in Figure.2 and the simulation result of electric field distribution is shown in Figure.3. III. RESULTE The voltage of Phase A is set to kV, according to the phase relations, the other two phase voltage is set to . The position of UAV is expressed by D and H. D is the distance between the UAV and the transmission line"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003232_detc2016-59194-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003232_detc2016-59194-Figure5-1.png",
        "caption": "Fig. 5 Motion modes of the multi-mode 7R mechanism.",
        "texts": [
            " The two configurations associated with J9 belong to the motion mode 3 and there is no new motion mode associated with J9. Since there is no real solution of \u03b83 to the equation u3 + 4 = 0, there is no motion mode associated with the irreducible component J12. 5.1.1 Planar 4R mode The 1-dimensional vanishing set of V(J )14 of J14 leads to u2 \u2212 1 = 0 u4 \u2212 1 = 0 u6 \u2212 1 = 0 \u00b7 \u00b7 \u00b7 (45) i.e.,  \u03b82 = 0 \u03b84 = 0 \u03b86 = 0 \u00b7 \u00b7 \u00b7 (46) Equation (46) shows that the motion mode associated with V(J )14 is a 1-DOF planar 4R mode (Fig. 5(a)). In this mode, joints 2, 4, and 6 lose their DOF and the joint axes of joints 1, 3, 5 and 7 are parallel. The input-output equation of the mechanism in this mode can be represented using the equations that were omitted in Eq. (46). 5.1.2 Orthogonal Bricard 6R mode The 1- dimensional vanishing set of V(J )1 of J1 leads to \u22121 + u7 = 0 u1 \u2212 u3 = 0 u4 \u2212 u2 = 0 u5 \u2212 u3 = 0 u6 \u2212 u2 = 0 \u00b7 \u00b7 \u00b7 (47) 7 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90696/ on 01/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use i.e.,  \u03b87 = 0 \u03b81 = \u03b83 = \u03b85 \u03b84 = \u03b82 = \u03b86 \u00b7 \u00b7 \u00b7 (48) Equation (48) shows that the motion mode associated with V(J )1 is a 1-DOF orthogonal Bricard 6R mode (Fig. 5(b)). In this mode, joint 7 loses its DOF and the joint axes of joints 1 and 6 are perpendicular to each other. The input-output equation of the mechanism in this mode can be represented using the equations that were omitted in Eq. (48). 5.1.3 Plane symmetric 6R mode The 1- dimensional vanishing set of V(J )13 of J13 leads to 1 + u1 = 0 u5 \u2212 u3 = 0 u6 \u2212 u2 = 0 \u00b7 \u00b7 \u00b7 (49) i.e.,  \u03b81 = \u03c0 \u03b85 = \u03b83 \u03b86 = \u03b82 \u00b7 \u00b7 \u00b7 (50) Equation (50) shows that the motion mode associated with V(J )13 is a 1-DOF plane symmetric 6R mode (Fig. 5(c)). In this mode, joint 1 loses its DOF and the distance between the joint axes of joints 2 and 7 is equal to the link length of link 6. The input-output equations of the mechanism in this mode can be represented using the equations that were omitted in Eq. (50). In this section, transition configurations between each pair of modes of the multi-mode 7R mechanism will be identified. The transition configuration between two motion modes can be obtained using the set of the equations composed of both set of equations associated with these motion modes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002832_978-981-10-1956-2_5-Figure5.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002832_978-981-10-1956-2_5-Figure5.10-1.png",
        "caption": "Fig. 5.10 Schematic diagram of the USV",
        "texts": [
            " Indeed, several research articles have proposed in the last few years to address the stabilization problem of USV, yet none of them could ensure ultimate performance for the system. Only they have achieved a tradeoff between different performance requirements, and thereby designing a control law for the USV is still being considered as an open research problem in the literature of nonlinear control engineering. Therefore, authors have also considered the stabilization problem of USV for demonstrating the versatility of the proposed control approach. The schematic diagram of the USV is shown in Fig. 5.10. The dynamic equation of the USV is described in Eq. (5.35). With the following choice of state variables: x1 \u00bc x; x2 \u00bc h; x3 \u00bc y, x4 \u00bc vx; x5 \u00bc x; x6 \u00bc vy. Consequently, nonlinear state equation of the USV has taken the structure of state model as described in the following Eq. (5.35) _x1 \u00bc x4 _x2 \u00bc x5 _x3 \u00bc x6 _x4 \u00bc u1 _x5 \u00bc u2 _x6 \u00bc u1 tan x2 \u00fe cy m x4 tan x2 x6\u00f0 \u00de \u00f05:35\u00de In the above representation, x1 and x2 represent the actuated variable, where x3 represents unactuated variable. Similarly, x4 and x5 represent the actuated velocity components, whereas x6 is the unactuated velocity components"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.20-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.20-1.png",
        "caption": "Fig. 3.20 Sample problem beam under constant distributed load and different boundary conditions: a cantilevered and b simply supported beam",
        "texts": [
            "9 as: ue z(x) = N2u(x)u2z + N2\u03d5(x)\u03d52y = \u23a1 \u23a33 ( x L )2 \u2212 2 ( x L )3 \u23a4 \u23a6 ( ML2 2EIz ) + [ x2 L \u2212 x3 L2 ]( \u2212ML EIz ) = Mx2 2EIy . (3.160) According to Ref. [37], this course matches with the analytical solution. Conclusion: Finite element solution and analytical solution are identical! 3.2 Sample: Beam under constant distributed load\u2014approximation through one single finite element Determine through one single finite element the displacement and the rotation (a) of the right-hand boundary and (b) in the middle for the beam under constant distributed load, which is illustrated in Fig. 3.20. Furthermore, determine the course of the bending line ue z = ue z(x) and compare the finite element solution with the analytical solution. 3.2 Solution To solve the problem, the constant distributed load has to be converted into equivalent nodal loads. These equivalent nodal loads can be extracted from Table3.7 for the considered case, and the finite element equation results to: 3.3 Finite Element Solution 131 EIy L3 \u23a1 \u23a2 \u23a2 \u23a3 12 \u22126L \u221212 \u22126L \u22126L 4L2 6L 2L2 \u221212 6L 12 6L \u22126L 2L2 6L 4L2 \u23a4 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a3 u1z \u03d51y u2z \u03d52y \u23a4 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 \u2212 qL 2 + qL2 12 \u2212 qL 2 \u2212 qL2 12 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 . (3.161) (a) Consideration of the boundary conditions shown in Fig. 3.20a, meaning the fixed support on the left-hand boundary, and solving for the unknowns yields: [ u2z \u03d52y ] = L 12EIz [ 4L2 \u22126L \u22126L 12 ][\u2212 qL 2 \u2212 qL2 12 ] = \u23a1 \u23a3 \u2212 qL4 8EIy + qL3 6EIy \u23a4 \u23a6 . (3.162) The analytical solution according to Ref. [37] yields uz(x = L) = \u2212 q 24EIy ( 6L4 \u2212 4L4 + L4 ) = \u2212 qL4 8EIy , (3.163) or alternatively the rotation based on \u03d5y(x) = \u2212 duz(x) dx : \u03d5y(x) = + q 24EIz ( 12L2x \u2212 12Lx2 + 4x3 ) (3.164) or only at the right-hand boundary: \u03d5y(x = L) = + q 24EIy ( 12L3 \u2212 12L3 + 4L3 ) = + qL3 6EIy ",
            " [37] results in uz(x) = \u2212 q 24EIy( x4 \u2212 4Lx3 + 6L2x2 ) , meaning the analytical and therefore the exact course is not identical with the numerical solution between the nodes (0 < x < L), see Fig. 3.21. One can see that between the nodes a small difference between the two solutions arises. If a higher accuracy is demanded between those two nodes, the beam has to be divided into more elements. 132 3 Euler\u2013Bernoulli Beams and Frames Fig. 3.21 Comparison of the analytical and the finite element solution for the beam according to Fig. 3.20a Conclusion: Finite element solution and the analytical solution are only identical at the nodes! (b)Consideration of the boundary conditions shown inFig. 3.20b,meaning the simple support and the roller support, yields through the elimination of the first and third line and column of the system of Eq. (3.161): EIy L3 [ 4L2 2L2 2L2 4L2 ] [ \u03d51y \u03d52y ] = [ + qL2 12 \u2212 qL2 12 ] . (3.167) Solving for the unknowns yields: [ \u03d51y \u03d52y ] = 1 12EIzL [ 4L2 \u22122L2 \u22122L2 4L2 ][+ qL2 12 \u2212 qL2 12 ] = [+ qL3 24EIy \u2212 qL3 24EIy ] . (3.168) The course of the bending line ue z = ue z(x) results from Table3.9 as: ue z(x) = N1\u03d5(x)\u03d51y + N2\u03d5(x)\u03d52y = [ \u2212x + 2 x2 L \u2212 x3 L2 ]( + qL3 24EIy ) + [ +x2 L \u2212 x3 L2 ]( \u2212 qL3 24EIy ) = \u2212 q 24EIy (\u2212L2x2 + L3x ) , (3.169) however the analytical course according to Ref. [37] results in uz(x) = \u2212 q 24EIy( x4 \u2212 2Lx3 + L3x ) , meaning the analytical and therefore exact course is also at this point not identical with the numerical solution between the nodes (0 < x < L), see Fig. 3.22. 3.3 Finite Element Solution 133 Fig. 3.22 Comparison of the analytical and the finite element solution for the beam according to Fig. 3.20b The numerical solution for the deflection in the middle of the beam yields ue z(x = 1 2L) = \u22124qL4 384EIy , however the exact solution is uz(x = 1 2L) = \u22125qL4 384EIy . Conclusion: Finite element solution and analytical solution are only identical at the nodes! 3.3 Example: Beam with distributed load over half of the length The following Fig. 3.23 shows a horizontal beam structure of length 2L which is fixed at both ends. The left-hand part of the structure (0 \u2264 X \u2264 L) is loaded by a constant distributed load q"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000832_ijcnn.2014.6889406-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000832_ijcnn.2014.6889406-Figure1-1.png",
        "caption": "Fig. 1. Mechanic model of the four-rotor helicopter.",
        "texts": [
            " Section 3 details the design of the disturbance observer. The adaptive scheme for updating the controller parameters and its stability analysis are developed in Section 4. The simulation results are presented in Section 5, followed by some conclusions in Section 6. II. Description of the System Model and Control Problems statement Four-rotor helicopter is an underactuated, dynamic vehicle with four input forces and 6 DOF motion. The helicopter has four propellers installed in a cross configuration as shown in Fig. 1. Ignoring the gyroscopic effect, bearing friction and atmospheric disturbance on the propellers, a simple model consists of three differential equations can be get based on the force condition of the system,. When a positive voltage is applied to a motor, a positive thrust is generated and this causes the corresponding propeller assembly to rise. The thrust force generated by the front, back, left and right propellers is denoted as fV , bV , lV and rV respectively. The thrust force generated by the front and back motors primarily actuate motions about the pitch axis while the right and left motors primarily move the hover about its roll axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000556_20140824-6-za-1003.01018-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000556_20140824-6-za-1003.01018-Figure2-1.png",
        "caption": "Fig. 2. General diagram of the virtual robot, where links L\u22121 . . . Ln represent the total links of the virtual robot, and L1 . . . Ln are the articulated object\u2019s links.",
        "texts": [
            " The information in W is enough to analyze whether G allows a FC grasp or not. For a planar object, 4 frictionless contacts are sufficient to assure the FC condition, i.e. a set of pointsG = {p1, ...,p4} allowing a set of wrenches W = {w1, . . . ,w4}. The generalized wrench for a serial articulated object described in this section is the generalization to n links of the procedure developed for 2 and 3 links by Alvarado and Sua\u0301rez (2013). The procedure considers a virtual robot of n+ 2 joints (see Fig 2) wherein the first and second joint are virtual ones and the rest of them are equivalent to the articulated object to be grasped.The following basic nomenclature is used in the procedure: Li: Link i of the virtual robot, i = \u22121, . . . , n, L\u22121 and L0 are virtual ones, while L1 to Ln are the real ones. qi: Joint i of the virtual robot, i = \u22122, . . . , n \u2212 1, q\u22122 to q0 are virtual ones, and q1 to qn\u22121 are the real ones. Qi: Position of the joint qi, i = 0, . . . , n\u22121, and position of the final end of the link Ln(i = n) respect to the frame base"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure5-1.png",
        "caption": "Figure 5 Description of the relationship between pitch deviation and cutter rotation angle",
        "texts": [
            " Therefore, spiral cutting marks are formed on the tooth flanks, and its pitch increases in proportion to the feed rate, which deteriorates the gear accuracy as shown in Fig. 4. However, because internal gear and cutter have a high contact ratio, cutter with pitch deviations, directly influences pitch deviations of the skived gear. Now let\u2019s focus in one specific tooth space of a gear. Pitch deviation is the difference of distance between adjacent tooth flank. Because in our model, simulations are conducted in only one tooth space of the workpiece, pitch deviation can be described as a rotational deviation of the cutter (Fig. 5). Furthermore, considering the fact that pitch deviation can be both positive and negative values, and that it is a kind of manufacturing error, one can assume that it follows a normal distribution of mean zero. Therefore, in the model which is explained later, pitch deviation was represented as the cutter rotation deviation that follows normal distribution of mean zero. In an actual skiving, the cutter shaft is long as 30mm while the run out of the cutter is small as 20 m. Therefore, the posture variation of the cutting edges can be ignored"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001149_2014-01-1797-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001149_2014-01-1797-Figure1-1.png",
        "caption": "Figure 1. Architecture of hybrid vehicle",
        "texts": [
            " Owing to these breakthrough advances, both fuel economy and acceleration time was improved from the first-generation system to achieve the fastest acceleration of all vehicles in its category. CITATION: Otokawa, K., Hayasaki, K., Abe, T., and Gunji, K., \"Performance Evolution of a One-motor Two-Clutch Parallel Full Hybrid System,\" SAE Int. J. Engines 7(3):2014, doi:10.4271/2014-01-1797. 1555 The architecture of the second-generation hybrid system is basically the same as that of the first-generation system, as shown in Figure 1 and Figure 2. The 1M2CL parallel full hybrid system is composed of gasoline engine, one electric motor/ generator, inverter, lithium-ion battery and two mechanical clutches. The motor can be used for several purposes-motoronly driving, engine assist, engine start, regeneration during coasting, and engine-based generation-owing to both the high-output, fast charging/discharging lithium-ion battery and the precise, high-speed motor control. As shown in Figure 3, the torque converter was removed from the transmission and a dry single-plate clutch (clutch-1) with a hydraulic actuator was positioned in its place to achieve both excellent fuel efficiency and a direct acceleration feeling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002529_s106345411602014x-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002529_s106345411602014x-Figure7-1.png",
        "caption": "Fig. 7. Control force for a trolley with two pendulums (case of m2/m1 = 1/64).",
        "texts": [
            " The sought-for functions are proportional to the trolley movement S in time , so this movement expressed in fractions of length of the longest pendulum was chosen so that the pendulums rotation angles would not exceed ten degrees. In Figs. 2\u20137, the solid lines correspond to results obtained with the new approach, and the dashed ones correspond to the results of the old method. Figures 2 and 3 correspond to the calculations when the trolley hosts one pendulum; the system parameters were set as follows: T T VESTNIK ST. PETERSBURG UNIVERSITY. MATHEMATICS Vol. 49 No. 2 2016 A NEW APPROACH TO FINDING 189 190 VESTNIK ST. PETERSBURG UNIVERSITY. MATHEMATICS Vol. 49 No. 2 2016 ZEGZHDA et al. and for Fig. 7 The calculations predictably showed that the bigger the mass of the trolley compared to the pendulums, the closer the results of the first and the second approach. Therefore the trolley mass was selected to be half the mass of the first main pendulum. The length and the mass of the second pendulum in relation to the first one were selected to be 1/4 and 1/8, respectively. The trolley movement is equal to 0.2l1, where l1 is the length of the main pendulum. We see that application of the new approach results in reduction of the pendulums oscillations both in amplitude and in frequency"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.6-1.png",
        "caption": "Fig. 10.6 Approximation of the B\u2013H characteristics of a magnetic core material: (a) by neglecting both hysteresis and saturation, (b) by neglecting hysteresis",
        "texts": [
            "9) The quantity \u03bc0 is the permeability of free space, and is equal to 4\u03c0 \u00b7 10\u22127 Henries per meter in MKS units. Figure 10.5b illustrates the B\u2013H characteristic of a typical iron alloy under highlevel sinusoidal steady-state excitation. The characteristic is highly nonlinear, and exhibits both hysteresis and saturation. The exact shape of the characteristic is dependent on the excitation, and is difficult to predict for arbitrary waveforms. For purposes of analysis, the core material characteristic of Fig. 10.5b is usually modeled by the linear or piecewise-linear characteristics of Fig. 10.6. In Fig. 10.6a, hysteresis and saturation are ignored. The B\u2013H characteristic is then given by B = \u03bcH \u03bc = \u03bcr\u03bc0 (10.10) The core material permeability \u03bc can be expressed as the product of the relative permeability \u03bcr and of \u03bc0. Typical values of \u03bcr lie in the range 103 to 105. The piecewise-linear model of Fig. 10.6b accounts for saturation but not hysteresis. The core material saturates when the magnitude of the flux density B exceeds the saturation flux density Bsat. For | B | < Bsat, the characteristic follows Eq. (10.10). When | B | > Bsat, the model predicts that the core reverts to free space, with a characteristic having a much smaller slope approximately equal to \u03bc0. Square-loop materials exhibit this type of abrupt-saturation characteristic, and additionally have a very large relative permeability \u03bcr",
            " Since there are n turns of wire passing through the window, each carrying current i(t), the net current passing through the window is ni(t). Hence, Ampere\u2019s law states that H(t) m = ni(t) (10.14) Let us model the core material characteristics by neglecting hysteresis but accounting for saturation, as follows: B = \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 Bsat for H \u2265 Bsat/\u03bc \u03bcH for |H| < Bsat/\u03bc \u2212Bsat for H \u2264 \u2212Bsat/\u03bc (10.15) The B\u2013H characteristic saturated slope \u03bc0 is much smaller than \u03bc, and is ignored here. A characteristic similar to Fig. 10.6b is obtained. The current magnitude Isat at the onset of saturation can be found by substitution of H = Bsat/\u03bc into Eq. (10.14). The result is Isat = Bsat m \u03bcn (10.16) We can now eliminate B and H from Eqs. (10.13) to (10.15), and solve for the electrical terminal characteristics. For |I| < Isat, B = \u03bcH. Equation (10.13) then becomes v(t) = \u03bcnAc dH(t) dt (10.17) Substitution of Eq. (10.14) into Eq. (10.17) to eliminate H(t) then leads to v(t) = \u03bcn2Ac m di(t) dt (10.18) which is of the form v(t) = L di(t) dt (10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.16-1.png",
        "caption": "FIGURE 6.16",
        "texts": [
            "15 uses a detailed \u2018as is\u2019 approach representing each of the leaves as a series of distributed lumped masses interconnected by beam elements with the correct sectional properties for the leaf. This type of model is also complicated by the need to model the interleaf contact forces between the lumped masses with any associated components of sliding friction. Assembling such models is greatly eased by the presence of some kind of macro; some software toolkits offer such macros but others leave the user to devise it. As shown in Figure 6.16 anti-roll bars may be modelled using two parts connected to the vehicle body by revolute joints and connected to each other by a torsional spring located on the centre line of the vehicle. In a more detailed model the analyst could include rubber bush elements rather than the revolute joints shown to connect each side of the anti-roll bar to the vehicle. In this case for a cylindrical bush the torsional stiffness of the bush would be zero to allow rotation about the axis, or could have a value associated with the friction in the joint"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001380_sami.2015.7061901-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001380_sami.2015.7061901-Figure5-1.png",
        "caption": "Figure 5 Orientation of test plates",
        "texts": [
            " Characteristics of the building platform The characteristics of the platform, or its dimensions and the material from which it is made, are exactly determined by the maker and seller of the specific device itself, or a machine using the DMLS technology for creating the product on the basis of the determined requirements and needs (Fig.4). C. Method of orientating the test plates on a platform After considering the conditions or rules for correct orientation of the model/product on the platform, several methods of orientation of the test plates were proposed (Fig. 5), whereby in the practical part (i.e. determining the roughness of the surface for the individual test plates in the individual sectors) of the given work attention is concentrated on the first method of orientation of the test plates (Fig. 5 - detail). Other methods of orientation are a question for possible future studies. D. Method of arranging the test plates on the platform The proposed method of arranging the test plates, with respect to the beginning of the coordinate system of the platform, derives from the assumption of changes of individual parameters (e.g. roughness), according to their mutual placement, with sufficient distance between the individual testing plates, or sectors, without a so-called \u201cinfluence zone\u201d. The arranging of the individual test plates into sectors labelled, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001238_978-90-481-9707-1_130-Figure12.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001238_978-90-481-9707-1_130-Figure12.5-1.png",
        "caption": "Fig. 12.5 How Dexterous Hexrotor moves: (a) hover, (b) roll, (c) pitch, (d) yaw, (e) translational acceleration along X axis, and (f) translational acceleration along Y axis",
        "texts": [
            " Therefore in-plane components result while still maintaining a symmetric basis of vectors, and forces and torques around each axis can be produced. Since Dexterous Hexrotor can span the force/torque space, by detaching and combining forces and torques produced by each rotor, forces and torques acting on the UAV around each axis can be accomplished, and controllability over full 6 degrees of freedom can be achieved. By varying the speed and choosing the direction of the rotation of the rotors, Dexterous Hexrotor can get not only forces along Z axis and torques around X, Y, Z axes but also forces along X, Y axes. As shown in Fig. 12.5, with the force along Z axis and torques around X, Y, Z axes, Dexterous Hexrotor can hover, roll, pitch, and yaw just like typical quadrotor does. But instead of pitching or rolling an angle like typical quadrotors, Dexterous Hexrotor can get translational acceleration by simply varying speed of these rotors. It can truly control its 6 degrees of freedom mobility. There are many ways to use six motors to create six independent degrees of freedom, but to create a design that was easy to fabricate, a disk design was used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure12-1.png",
        "caption": "Figure 12 Schematic diagram of the cutter and workpiece in a skiving rig. The shaft angle is given by rotating the C axis.",
        "texts": [
            " In the profile of condition b), no characteristic period is observed and it seems to be random. The lead deviation curves of condition a) in the Figure 11, have the same period of 0.25 mm (same as feed rate) on both left and right flank. In the condition b) the period of around 1mm is observed on both left and right flank. The amplitude of the both deviation curve, obviously increases for condition b). In the experiment, a modified bevel-gear cutting machine which has six axis (C29 manufactured by Klingelnberg GmbH) was used. Figure 12 shows a layout of a cutter and a workpiece spindle axes of the machine. The cutter was fed by moving the cutter spindle along the x and y axes simultaneously. The rotation speed of the cutter and workpiece were controlled by AC servo motors and the cutter rotation correction equivalent to helix angle of the workpiece were synchronously controlled by a CN system supplied by Siemens (840 D) The cutter spindle rotation speed was set to 700 min-1, which is equivalent to a cutting velocity of 80 m/min, and the rotation direction was clockwise"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001911_gt2015-44068-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001911_gt2015-44068-Figure12-1.png",
        "caption": "Figure 12: Typical brush seal layout",
        "texts": [],
        "surrounding_texts": [
            "Bristle diameter is one of the most important design parameters, which directly affects BTF and bristle stress levels at pressurized/unpressurized rotor interference conditions as well as steady state case. Typical industrial application for selecting the appropriate bristle diameter strongly depends on the experience, and larger diameters are usually preferred for applications that operate under high pressure loads. In order to improve the pressure load capacity of the brush seal, multi-stage and multi-layer configurations have been tried [18, 19, 20, 21, 22, 23]. However, due to nonlinear pressure drop between seal stages for multi-stage applications and introduced bristle locking phenomenon in multi-layer seal applications, the improvement in pressure load capacities are limited. Maximum pressure load capacity for a brush seal is reported as 27 bar (Dinc et al. [24]). In this section, the effect of bristle diameter on BTF and stress levels under transient and steady state conditions will be analyzed through correlated CAE models. Since the axial and tangential spacing parameters change with the bristle diameter, the number of bristle rows also changes (if other parameters such as bristle density, cant angle etc. are kept constant). Unpressurized rotor interference-BTF. The effect of bristle diameter on BTF at unpressurized-nonrotating rotor interference conditions (free-state rotor rub) has been examined by running the cases detailed in Tab. 4. Free-state BTF (BTF under unpressurized rotor rub) change with bristle diameter is given in Fig. 14 at different rotor interference levels. Second moment of area for the bristle crosssection is a 4th order function of the bristle diameter, therefore required tip force for bending the bristles is higher for larger bristle diameters. Increase in free-state BTF with bristle diameter is more pronounced at high interference levels. 7 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Steady state (\u0394P=0.3 MPa)\u2013VM Stress. The effect of bristle diameter on bristle stress levels for a loaded seal (pressure load without rotor interference) has been examined by running the cases detailed in Tab. 5. Stress levels at the most critical section, which is the FH point of backing plate side bristles (where downstream side bristles touching the backing plate corner, Figures 10 and 11), are examined during characterization study."
        ]
    },
    {
        "image_filename": "designv11_64_0001911_gt2015-44068-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001911_gt2015-44068-Figure4-1.png",
        "caption": "Figure 4: Brush seal CAE model-C3D8I, Pressure load profile",
        "texts": [
            " Rotor and backing plate have been modeled as rigid bodies, and backing plate corner radius has been introduced at FH region. The pressure profile for a loaded seal is extracted from the 7th order polynomial fit to experimental data (in seal radial direction) of Bayley et al. [13] and Turner et al. [14]. Linear pressure drop is used through the bristle pack in rotor axial direction (Almost linear pressure drop at fence height region has been reported in the literature [15, 16]). Pressure profile for brush seal FE models using C3D8I and B31 elements are given in Fig. 4 and Fig. 5 respectively. Bristles are modeled by using Haynes 25 material properties, which is typically preferred in most of turbine sealing applications due to their superior strength and satisfactory ductility (up to 600oC). Cold worked Haynes 25 material with 10% cold reduction has 725MPa tensile yield strength and 1070MPa ultimate tensile strength limits at room temperature [17]. Analyses have been conducted by using two CAE models, one of which has bristles with C3D8I elements and the other one has B31 bristles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002452_ijrapidm.2015.073549-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002452_ijrapidm.2015.073549-Figure11-1.png",
        "caption": "Figure 11 Loading conditions of the beam and the coordinate systems for different specimen configurations (see online version for colours)",
        "texts": [
            " The deflection of all the parts at 200N was picked as the point of comparison among experimental, numerical, and analytical results. 200N was chosen because it was approximately the mid-point on the linear part of elastic curve for all graphs. The average deflections of all the specimen configurations at 200 N and the distribution are shown in Table 3. The FEA simulations of the 3-point bending test were performed using ABAQUS. The specimen was modelled as a deformable body and was regarded as linear elastic orthotropic material with the properties shown in equation (6). As depicted in Figure 11, each specimen configuration was modelled with the corresponding material orientation, e.g. four coordinate systems were assigned to four specimen configurations 3P1, 3P2, 3P3, and 3P4. An additional rotation of 45\u00b0 about the z-axis was made to coincide the global coordinate system of ABAQUS with the machine coordinate system. Specimens were prescribed with the same boundary condition with the right and left bottom line fixed. An evenly distributed load of total 200N was applied on all the nodes along the loading line"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003137_0954410016676847-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003137_0954410016676847-Figure9-1.png",
        "caption": "Figure 9. One case of reducing the searching interval for the cubic Be\u0301zier curve algorithm (solid line: original curve; dashed line: cubic Be\u0301zier curve).",
        "texts": [
            "0\u00f01 ei\u00de 3 \u00fe 3!1\u00f01 ei \u00de 2 ei \u00fe 3!2\u00f01 ei\u00de 2 ei \u00fe !3 3 ei , i \u00bc 1, 2, 3 \u00f0100\u00de and the corresponding residuals in the root-finding problem are zei \u00bc F\u00f0!ei\u00de, i \u00bc 1, 2, 3 \u00f0101\u00de Among the three possibilities, the estimated root !e is selected as !e \u00bc !ei , if !ei 2 \u00f0!min,!max\u00de and jzei j \u00bc min\u00f0jze1 j, jze2 j, jze3 j\u00de, i \u00bc 1, 2, 3 \u00f0102\u00de If the estimated root !e is not accurate, the iterative algorithm computes the estimated root with a new searching range until the root is accurate enough. As illustrated in Figure 9, denoting Pe \u00bc \u00f0!e,F\u00f0!e\u00de\u00de, P1 \u00bc \u00f0!1,F\u00f0!1\u00de\u00de and P2 \u00bc \u00f0!2,F\u00f0!2\u00de\u00de, new endpointseP0 \u00bc \u00f0 ~!0, ~z0\u00de and eP3 \u00bc \u00f0 ~!3, ~z3\u00de can be selected from points P0, P1, P2, Pe, Pm and P3 with the smallest searching interval and ~z0 ~z3 5 0. Root-finding algorithm using a non-rational quadratic Be\u0301zier function. Quadratic Be\u0301zier curves are employed to model the convex variations of z and ! in the interval ! 2 [!0, !2]. For a certain iterative process, the quadratic Be\u0301zier curve can be formed as !\u00f0 \u00de \u00bc !0\u00f01 \u00de 2 \u00fe 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure8.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure8.3-1.png",
        "caption": "Fig. 8.3 Diagram illustrating the diffraction-limited spot size",
        "texts": [
            "2 mm and 13 mm, though only smaller sizes are used for welding. The depth of penetration is proportional to the amount of power supplied. Such depth is also dependent on the location of the focal point; penetration is maximized when the focal point is slightly below the surface of the workpiece. Millisecond-long pulses are used to weld thin materials, while continuous laser systems are employed for deep welds. In addition, a beam of finite diameter is focused by a lens onto a plate with a transverse electromagnetic mode pattern \u201c000\u201d as shown in Fig. 8.3. The individual parts of the beam striking the lens can be imagined as point radiators of a new wavefront. The lens will draw the rays together at the focal plane, and constructive or destructive interference will take place there. When two rays arrive at the screen and are half a wavelength (denoted as \u03bb) out of phase, then they will destructively interfere with a minimal light intensity. Thus, if the ray AB (Fig. 8.3) is \u03bb/2 longer than ray CB, the point B will represent the first dark ring. The central maximum will contain approximately 86% of the total beam power. The radius of this central maximum will be the focused beam radius (Rbeam). Considering the initial beam radius before focused is Rl, the distance of AB longer than CB would be \u03bb/2, which can be related to the lens radius as \u03bb \u00bc 2Rlsin\u03c6, where \u03c6 is the angle of ABC. Besides, the beam radius Rbeam can be related to the focal length of lens fl as Rbeam \u00bc fltan\u03c6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001377_s00542-014-2085-z-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001377_s00542-014-2085-z-Figure6-1.png",
        "caption": "Fig. 6 Temperature measurement configuration and setup",
        "texts": [
            " (1995) utilizing polysilicon piezoresistors fabricated on the silicon surface through a complex process; nevertheless, the substrate temperature with respect to the etching pressure was not discussed. Since the etching process was performed in a vacuum chamber filled with XeF2 gas, a non-contact infrared thermometer with a detecting wavelength between 1.95 and 2.5 \u03bcm (FTK9S-r80a-10S61, Japan Sensor corp.) was adopted because wavelengths above 3 \u03bcm, which are used in conventional infrared thermometers, cannot penetrate the common glass of the reaction chamber. The experimental configuration and setup are shown in Fig. 6. The specifications of the thermometer are listed in Table 2. The silicon surface is used for the temperature measurements since, firstly, the TFPM is very small (\u03a63 mm \u00d7 14.8 \u03bcm) compared with the silicon substrate (20 mm \u00d7 20 mm \u00d7 200 \u03bcm) and thus can be neglected, and secondly, the PDMS has very small thermal conductivity [0.18 W/(m K)] compared to silicon [130 W/(m K)], thus can be regarded, like the vacuum, as a thermal insulator. The wavelength of the chosen infrared thermometer is from 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003874_b978-0-12-396502-8.00002-4-Figure2.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003874_b978-0-12-396502-8.00002-4-Figure2.8-1.png",
        "caption": "FIGURE 2.8",
        "texts": [
            " Also, when we find non-distinct roots of the characteristic equation (assume multiplicity m as an example), the corresponding term in yP (t) shall be multiplied by tm . An example of a second order ordinary differential equation (ODE) with constant coefficients is considered as follows: LC d2vo(t) dt2 + RC dvo(t) dt + vo(t) = vi (t), (2.29) where R = 2.0 (in ), L = 2.0 (in H), C = 1/2 (in F), and vi (t) = e\u2212t (in V). This equation describes the input \u00d7 output relationship of the electrical circuit shown in Figure 2.8 for t > 0. Replacing the values of the components and the input voltage, the ODE becomes d2vo(t) dt2 + dvo(t) dt + vo(t) = e\u2212t\ufe38\ufe37\ufe37\ufe38 f (t) . (2.30) The associated characteristic equation s2 + s + 1 = 0 has roots s = \u22121\u00b1 j \u221a 3 2 leading to an homoge- neous solution given by vo H (t) = k1 cos (\u221a 3 2 t ) e\u2212t/2 + k2 sin (\u221a 3 2 t ) e\u2212t/2 = k3 cos (\u221a 3 2 t + k4 ) e\u2212t/2. (2.31) Next, from the forcing function f (t) = e\u2212t , we assume vo P (t) = k5e\u2212t and replace it in (2.30), resulting in k5 = 1 such that: vo(t) = k1 cos (\u221a 3 2 t ) e\u2212t/2 + k2 sin (\u221a 3 2 t ) e\u2212t/2 + e\u2212t = k3 cos (\u221a 3t 2 + k4 ) e\u2212t/2 + e\u2212t , (2",
            " Therefore, since the voltage across C and the current through L do not alter instantaneously, we know that vo(0+) = 1 and i(0+) = 0. Since we know that i(t) = C dvo(t) dt , we have dv0(t) dt \u2223\u2223\u2223 t=0+ = 0. With these initial conditions and from (2.32), we find k1 = 0 and k2 = 2 \u221a 3 3 . Finally, the general solution is given as vo(t) = 1.1547 sin (0.866t)e\u2212t/2 + e\u2212t . (2.33) Figure 2.9 shows vo(t) for 0 \u2264 t \u2264 10. An easy way to obtain this result with Matlab\u00a9 is: > y=dsolve (\u2019D2y+Dy+y=exp(-t)\u2019,\u2019y(0)=1\u2019,\u2019Dy(0)=0\u2019); > ezplot (y,[0 10]) Output voltage as a function of time, vo(t), for the circuit in Figure 2.8. See Video 3 to watch animation. To end this section, we represent this example using the state-space approach. We first rewrite (2.29) using the first and the second derivatives of vo(t) as v\u0307 and v\u0308, respectively, and also u = vi (t) as in LC v\u0308 + RC v\u0307 + v = u. (2.34) In this representation, we define a state vector x = [v v\u0307]T and its derivative x\u0307 = [v\u0307 v\u0308]T , and from these definitions we write the state equation and the output equation:\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 x\u0307 = Ax + Bu = [ 0 1 \u22121 LC \u2212R L ] [ v v\u0307 ] + [ 0 1 LC ] vi (t), y = Cx + Du = [10] [ v v\u0307 ] + 0vi (t), (2",
            "14): ( ) Equivalent circuit of an inductor in the Laplace domain. Equivalent circuit of a capacitor in the Laplace domain. with I (s) = X(s) R + 1 sC . (2.69) Substituting (2.69) in (2.68) and applying the inverse Laplace transform (Table 2.2) and and the linearity property (Table 2.3), we get L\u22121{Y (s)} = L\u22121 { 1 s \u2212 1 s + 1 RC } , y(t) = u(t)\u2212 e\u2212t 1 RC u(t), (2.70) which is the same result presented in (2.5). Another example, where the Laplace transform is useful, is the RLC circuit displayed by Figure 2.8 where the desired response is the voltage across the capacitor C, vo(t) = vC (t). Note that the initial conditions are part of the transform, as well as the transient and steady-state responses. Given the input signal vi (t) = e\u2212t u(t) with initial conditions vo(0\u2212) = 1, v\u0307o(0\u2212) = 0, and io(0\u2212) = 0, applying the Kirchhoff\u2019s voltage law to the circuit, we have: vi (t) = vR(t)+ vL(t)+ vC (t). (2.71) Directly substituting (2.62), (2.63) and (2.65) in (2.71), and applying the Laplace transform to the input signal, vi (t) = e\u2212t u(t) , we get 1 s + 1 = ( RI (s)+ sL I (s)\u2212 Lio(0 \u2212)+ 1 sC I (s)+ vo(0\u2212) s ) = I (s) ( R + sL + 1 sC ) \u2212 Lio(0 \u2212)+ vo(0\u2212) s = I (s) sC (s2LC + s RC + 1)\u2212 Lio(0 \u2212)+ vo(0\u2212) s , (2",
            " If the power associated with a (complex) signal x(t) is |x(t)|2, Parseval\u2019s Relationship states that the signal is represented equivalently in either the time- or frequency-domain without lost or gain of energy:\u222b \u221e \u2212\u221e |x(t)|2dt = 1 2\u03c0 \u222b \u221e \u2212\u221e |X( j )|2d . (2.124) Hence, we can compute average power in either the time- or frequency-domain. The term |X( j )|2 is known as the energy spectrum or energy density spectrum of the signal and shows how the energy of x(t) is distributed across the spectrum. A list of Fourier transform properties are summarized in Table 2.6. One more time, we take, as an example, the RLC circuit displayed by Figure 2.8, where the desired response is the voltage, vo(t), across the capacitor C, vC (t). If we use the Fourier transform, assisted by its differentiation property (Table 2.6), to solve the ODE in (2.30) we have F{e\u2212t } = Vi ( j ) = F { LC d2vo(t) dt2 + RC dvo(t) dt + vo(t) } (2.125) and we get the following result Vi ( j ) = \u2212LC 2Vo( j )+ j RC Vo( j )+ Vo( j ), (2.126) where the output in the frequency-domain, Vo( j ), is given by Vo( j ) = Vi ( j ) (1\u2212 LC 2)+ j RC , (2.127) which exhibits a resonant behavior exactly as in (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure8-1.png",
        "caption": "Figure 8. (a) Example 2: System under consideration, and (b) angles and unit vectors defining positions of the springs.",
        "texts": [
            "2 as a dotted line, and the value of A is taken arbitrarily to be equal to unity. It is important to note that the angles between these two directions and the horizontal are 75o and 15o, which agrees with the position of two-element equivalent system of springs shown in Figure 6. However, unlike this procedure related to natural modes of vibration, which involves lengthy calculations, the one related to the equivalent system of springs performed previously is straightforward and mathematically tractable. Example 2. The second example is shown and defined in Figure 8(a). The system consists of four springs, whose stiffness coefficients are k1\u00bc 4000N/m, k2\u00bc 2000N/m, k3\u00bc 2500N/m and k4\u00bc 1184.6N/m. They are joined in a point M0, where the particle of mass m (mg\u00bc 900N) is located. In the static equilibrium position shown, the first and the third spring are deformed in this position, where 1\u00bc 0.18028m and 3\u00bc 0.2m (this implies that 2\u00bc 0m and 4\u00bc 0m). Note that it is easy to check that the system of all forces acting at point M0 is balanced. In addition, based on Figure 8(a), one can find the lengths of the springs in the static equilibrium position l1\u00bc ffiffiffiffiffi 13 p m, l2\u00bc 3m, l3\u00bc 5m and l4\u00bc 6m. The position of each of the spring is determined based on Figure 8(a) and (b) as \u20191\u00bc 3 / 2\u00feb, \u20192\u00bc 3 /2, \u20193\u00bc 3 /2\u2013g and \u20194\u00bc , where \u00bc arctan 2=3\u00f0 \u00de \u00bc 0:588 and \u00bc arctan 4=3\u00f0 \u00de \u00bc 0:927. The angle k from equation (1) is defined by tan 2 k \u00bc P4 i\u00bc1 ki 1 i li sin 2\u2019iP4 i\u00bc1 ki 1 i li cos 2\u2019i \u00bc 0:75 \u00f042\u00de which gives k\u00bc 0.322\u00bc 18.435o. Equation (2a,b) yields the stiffness coefficients of the equivalent system plotted in Figure 9(a) kI \u00bc X4 i\u00bc1 ki cos2 \u2019i k\u00f0 \u00de \u00fe i li sin2 \u2019i k\u00f0 \u00de \u00bc 3989:2 N=m, kII \u00bc X4 i\u00bc1 ki sin2 \u2019i k\u00f0 \u00de \u00fe i li cos2 \u2019i k\u00f0 \u00de \u00bc 5995:4 N=m \u00f043\u00de Note that this equivalent system does not contain the static force mg and the two new springs are not pre-stressed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure1.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure1.2-1.png",
        "caption": "Fig. 1.2 Airbus A380 a geometry and b finite element mesh",
        "texts": [],
        "surrounding_texts": [
            "2 1 Introduction to the Finite Element Method\nFigures1.2 and 1.3 illustrate this process where first a geometrical representation of an engineering structure is shown. In a second step, this geometry is approximated based on smaller geometrical entities, so-called finite elements.\nIn the context of engineering education, it must be stated that courses on the finite element method require a certain foundation (see Fig. 1.4) which are normally provided during the first years of study. This may imply that students face some difficulties compared to the early foundation courses because the comprehensive treatment of thismethod assembles a considerable amount of engineering knowledge.\nEngineers describe physical phenomena and processes typically by equations, particulary by partial differential equations [15, 18, 46]. In this context, the derivation and the solution of these differential equations (see Fig. 1.5) is the task of engineers, obviously requiring fundamental knowledge from physics and engineering mathematics.\nThe importance of partial differential equations is clearly represented in the following quote: \u2018For more than 250 years partial differential equations have been clearly the most important tool available to mankind in order to understand a large variety of phenomena, natural at first and then those originating from human activity and technological development. Mechanics, physics and their engineering applications were the first to benefit from the impact of partial differential equations on modeling and design,...\u2019 [20].\nIn the one-dimensional case, a physical problem can be generally described in a spatial domain \u03a9 by the differential equation",
            "1 Introduction to the Finite Element Method 3\nL{y(x)} = b (x \u2208 \u03a9) (1.1)\nand by the conditions which are prescribed on the boundary \u0393 . The differential equation is also called the strong form or the original statement of the problem. The expression \u2018strong form\u2019 comes from the fact that the differential equation describes exactly each point x in the domain of the problem. The operatorL{. . .} in Eq. (1.1) is an arbitrary differential operator which can take, for example, the following forms:",
            "4 1 Introduction to the Finite Element Method\nL{. . .} = d2\ndx2 {. . .}, (1.2)\nL{. . .} = d4\ndx4 {. . .}, (1.3)\nL{. . .} = d4 dx4 {. . .} + d dx {. . .} + {. . .}. (1.4)"
        ]
    },
    {
        "image_filename": "designv11_64_0001168_gt2014-26128-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001168_gt2014-26128-Figure5-1.png",
        "caption": "Figure 5: Image showing the secondary flow (azimuthal component removed) in the back chamber. Block arrows show the bulk direction of flow.",
        "texts": [
            " Within Fluent the Schiller and Naumann model [18] is used to define f as \ud835\udc53 = \ud835\udc36 \ud835\udc45\ud835\udc52 24 Eq 8 and \u03c4p is defined as \ud835\udf0f\ud835\udc5d = \ud835\udf0c \ud835\udc51 18\ud835\udf07 Eq 9 Eulerian Methodology Using a disperse phase approach, a droplet diameter must be set and in this study dp is chosen as 0.8 mm; this value was chosen with reference to the work of Glahn et al. [19]. For the same reasons as explained with regard to the VOF model, an explicit time-stepping approach was employed to satisfy C = 2. The same mesh as for the VOF simulation (containing 0.9 million cells) was employed. The computed air flow in the chamber is illustrated in Figure 5 where vectors of velocity magnitude are displayed. As can be seen, flow from the back of the gear is the driving force for the flows in this region. The jet curves up towards the top of the chamber, splitting on the outer rim and driving the chamber flow from front to back along the outer face and down the back wall towards the oil inlet. Towards the front of the chamber this flow contributes to a small but strong vortex cell at the shroud hole which then leaves through the hole as a jet. 5 Copyright \u00a9 2014 by Rolls-Royce plc Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000892_icrom.2014.6990948-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000892_icrom.2014.6990948-Figure4-1.png",
        "caption": "Fig. 4. The applied force and moment at the robot tip.",
        "texts": [
            " Coriolis and Centrifugal Effects Using Euler-Lagrange method of dynamics modeling, the derivatives of M represent the centrifugal and Coriolis effects. These terms can be represented by a vector V, which can be calculated by Christoffel symbols, as 1 2 6 3 6 3 6 3 * 1 1 [ ] , ( ) ( ) , T n n n i ijk j k V V V V m q j q k + + + = = = = \u2211\u2211 V (49) where q (j) is the jth element of q , and m* is defined as * ( , ) ( , ) . ( ) 2 ( )ijk i j j km q k q i \u2202 \u2202= \u2212 \u2202 \u2202 M M (50) C. Non-Conservative Work In this paper, the non-conservative force and moment are considered as a force and moment applied to the robot tip, as Ftip and \u03c4tip, as depicted in Fig. 4. By definition, the speed of work done by these force and moment is 1 1 ,T T n nW + += +\u03c4 \u03c9 F v (51) which can be resolved by substituting (35) and (36) as ( 1) ( 1) .T T i v iW \u03c9 + += +\u03c4 J q F J q (52) Thus, the general forces of the model can be represented by Q, as ( 1) ( 1) , . T T T i v i W \u03c9 + + = = + Q q Q J \u03c4 J F (53) D. Robot Model Finally, using the Euler-Lagrange method, the robot model is derived as ,e g+ + + =Mq V G G Q (54) where M is determined by (40), V is given by (49), Ge is from (45), Gg is defined in (48) and Q is yielded in (53)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000540_1754337115582121-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000540_1754337115582121-Figure2-1.png",
        "caption": "Figure 2. Free body diagram of the external forces acting on the rowing oar during the drive. Fh is the force applied by the rower to the handle, Fo is the normal reaction force at the oarlock, M is the resultant moment of force, Fb is the load on the blade, Lb is the beam moment arm, and Ls is the support moment arm.",
        "texts": [
            " The blades kinematics during the drive were reconstructed from the on-water measurements and compared with those of a perfectly rigid oar-shaft.14 Compared to the tested oars, the rigid oar-shaft assumption \u2018\u2018substantially\u2019\u2019 changed the reconstructed blade kinematics. The blade\u2019s path laid more toward the stern throughout the drive, which changed the hydrodynamic forces calculated on the blades.14 The effect of oar length on rowing biomechanics is also of interest. The external forces acting on the rowing oar during the drive are commonly illustrated using a lever model, as shown in Figure 2. Fh represents the effort applied by the rower to the handle, Fb is the load on the blade, and Fo is the normal reaction force at the oarlock, which is the sum of Fb and Fh. The lines of action are in the direction parallel to the boat\u2019s main motion (i.e. the x-axis). The support moment arm (Ls) is the perpendicular distance between the points of application of the force vectors Fh and Fo, and the beam moment arm (Lb) is the perpendicular distance between Fo and Fb. The moment of force about the oarlock (M) in dynamic equilibrium can be calculated via X M=FhLs FbLb Ia=0 \u00f03\u00de where a is the angular acceleration of the oar and I is mass moment of inertia"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure9.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure9.6-1.png",
        "caption": "Fig. 9.6 Transient analysis of a rod element: a schematic sketch of the problem; b force-time relationship",
        "texts": [
            "66) to solve for u\u03081: u1 = ( K + M \u03b2( t)2 )\u22121{ F1 + M \u03b2( t)2 ( u0 + t u\u03070 + ( 1 2 \u2212 \u03b2 ) ( t)2u\u03080 )} . 4 Use Eq. (9.67) to solve for u\u03081: u\u03081 = 1 \u03b2( )2 ( u1 \u2212 u0 \u2212 t u\u03070 \u2212 ( t)2 ( 1 2 \u2212 \u03b2 ) u\u03080 ) . 366 9 Integration Methods for Transient Problems 5 Use Eqs. (9.60) and (9.61) to solve for u\u03071: u\u03071 = u\u03070 + t ((1 \u2212 \u03b3)u\u03080 + \u03b3u\u03081). 6 Repeat steps 3\u20135 to obtain displacement, acceleration, and velocity for all other time steps. 9.3.4 Solved Problems 9.1 Example: Transient analysis of a rod element Given is a rod which is loaded by a time-dependent force Fx (t), see Fig. 9.6. Use a single rod element and the following initial conditions u(0) = u\u0307(0) = 0 to apply the solution scheme from p. 364. Further values are Fx (0) = 2000, t\u2217 = 0.2, E A L = 100, m = \u03b3 AL g = 90, t = 0.05. 9.1 Solution Let us start in the common manner, i.e. state the non-reduced system of equations. Since we consider only a single element: \u03b3 AL 6g [ 2 1 1 2 ] [ u\u03081x u\u03082x ] + E A L [ 1 \u22121 \u22121 1 ] [ u1x u2x ] = [\u2212R1 Fx(t) ] . (9.68) Consideration of the boundary condition at the left-hand end at x = 0, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002064_2016-01-1136-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002064_2016-01-1136-Figure9-1.png",
        "caption": "Figure 9. e-LSD actuator test-rig.",
        "texts": [
            " Since the torque variation between the starting point and the end of the simulation is about 3.5% of the maximum value, it is possible to consider only the asymptotic torque value in the vehicle dynamics simulations. The torque value in the graph is normalized on the maximum value and expressed as percentage. To exclude the possibility of interference between the eigenfrequencies of the e-LSD clutch and the frequency of the pressure signal imposed by the actuator, a dedicated test-rig, which is visible in the figure 9, was adopted. This is composed of a main frame where the right cartridge of the differential is housed. This is installed on a wood plate, where also the actuator and the electronic unit are hosted, as well as the oil tank. Please, note that the discs do not have a relative movement, being a static test-rig, because the main goal is to relate the oil pressure to the normal thrust on the clutch\u2019s discs. The process expected the actuator to generate a sinusoidal input of pressure to the differential clutch and the load cell installed in the middle of the frame to acquire the normal force of the clutch piston"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001503_20140824-6-za-1003.01713-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001503_20140824-6-za-1003.01713-Figure2-1.png",
        "caption": "Fig. 2. State of homogenous transformation of a MAS moving in a plane",
        "texts": [
            " Let ( ) denote position of a follower agent i (i=n+2, n+3, \u2026, N) at time , where the transient configuration of the MAS is a homogenous map of the initial formation, then, ( ) \u2211 ( ) ( ) It is noted that ( ) is the position of leader agent k (k=1, 2, \u2026, n+1) and s are the constant weights specified by eqn. (3) based on initial positions of follower agent i and leaders 1, 2, \u2026, n+1 i.e. the weight is denoted by where agents , , \u2026, and are all leader agents. Furthermore, is is at row i-n-1 and column k of ( ). It is noted that \u2211 ( ) As shown in Fig. 2, for the case of MAS evolution in a 2-D domain, three leader agents 1, 2, and 3 are located at the vertices of a triangle, called leading triangle, and followers are distributed inside the leading triangle. As seen, every follower is a disk with radius which is located inside a circular domain with radius called a safe zone. Our desire is that none of the follower agents leave the safe zone, throughout their evolution under local inter agent communication. In other words, deviation from state of homogenous transformation, \u2016 \u2016, is not larger than , or \u2016 \u2016 ( ) Furthermore, safe zone of every follower agent must not be penetrated by other agents"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001692_s00521-015-2117-3-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001692_s00521-015-2117-3-Figure1-1.png",
        "caption": "Fig. 1 Three topologies",
        "texts": [
            " h In this section, a numerical example is given to illustrate the feasibility and effectiveness of our theoretical results. A chaotic Chua\u2019s circuit model is considered as an isolated node of the dynamical network, which is described by: _xi\u00f0t\u00de \u00bc f \u00f0xi\u00f0t\u00de; t\u00de \u00bc \u00f010\u00f0xi2 g\u00f0xi1\u00de\u00de; xi1 xi2 \u00fe xi3; 18xi2\u00deT where xi\u00f0t\u00de \u00bc \u00f0xi1\u00f0t\u00de; xi2\u00f0t\u00de; xi3\u00f0t\u00de\u00deT with g\u00f0xi1\u00de \u00bc 1=4xi1 \u00fe1=2\u00f0 1=3 1=4\u00de\u00f0jxi1 \u00fe 1j jxi1 1j\u00de. Thus, by computation, we can get kKk \u00bc 4:3871 [30]. We assume that there are totally five nodes in network (1) and consider there are three random switching topolo- gies, which are shown in Fig. 1 with all the connectivity weights being 1. Let A \u00bc \u00f0aij\u00de 2 R5 5 with all off diagonal elements being 0.23, and diagonal elements being 0:92. Let b1 \u00bc 0:9; b2 \u00bc 0:6; b3 \u00bc 0:3; s\u00f0t\u00de \u00bc 0:2 0:1 sin\u00f0t\u00de, i.e., s \u00bc 0:1 and s \u00bc 0:3. Take M \u00bc 1;C \u00bc diag f0:9; 1; 0g; c1 \u00bc c2 \u00bc 40, by some calculations, we get d r\u00f0t\u00de max \u00bc d r\u00f0t\u00de min \u00bc a r\u00f0t\u00de max \u00bc a r\u00f0t\u00de min \u00bc 1; Gs \u00bc fg1; g2g; Gu \u00bc fg3g and c \u00bc 200. We thus have k1 \u00bc 4:1590; k2 \u00bc 5:1697; n3 \u00bc 13:3654. Determining a random switching signal r\u00f0t\u00de as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000318_tia.2010.2070784-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000318_tia.2010.2070784-Figure2-1.png",
        "caption": "Fig. 2. PM rotor topologies considered. (a) IPM rotor. (b) SPM rotor.",
        "texts": [
            " Referring to both interior PM (IPM) and surface-mounted PM (SPM) motor topologies and different geometrical arrangements of the two windings, the following items will be investigated: 1) the average torque and the torque ripple in various operating conditions (Section IV); 2) the unbalanced radial force in faulty operating conditions (Section V); 3) the mutual coupling between phases (Section VI); 4) the overload capability when only a winding is supplied (Section VII); 5) the short-circuit behavior (Section VII). As explained previously, this paper focuses on the analysis and design of the machine, neglecting the critical aspect of the control (i.e., the synchronization of the two supplies). Both PM motor topologies refer to a 12-slot 10-pole motor, characterized by a fractional-slot winding with nonoverlapped coils. The main data of the stator are reported in Table I. The IPM and SPM rotors are shown in Fig. 2. The nominal torque is 0093-9994/$26.00 \u00a9 2010 IEEE fixed to be 15 N \u00b7 m, and the nominal current is 12.4 A (peak). In order to achieve the required torque, both motors are supplied at a maximum torque per ampere point of the d\u2013q plane. The SPM motor requires a stack length that is equal to 50% of the IPM motor length. On the other hand, the IPM motor exhibits a higher flux-weakening capability, as computed analytically and shown in Figs. 3 and 4. In addition, even if the motor lengths are different, there is the same PM volume in the two rotors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001427_chicc.2014.6896462-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001427_chicc.2014.6896462-Figure1-1.png",
        "caption": "Fig. 1: Reference frames: Earth- xed and body- xed",
        "texts": [],
        "surrounding_texts": [
            "Some graph concepts are brie y introduced. A graph G = {V , E} consists of a node set V = {n1, ..., nN} and an edge set E = {(ni, nj) \u2208 V \u00d7 V} with (ni, nj) describes the communication from node i to node j. A directed path in the graph is an ordered sequence of nodes such that any two consecutive nodes in the sequence are an edge of the graph. A digraph has a spanning tree, if there is a node called as the root, such that there is a directed path from the root to every other node in the graph. De ne an adjacency matrix A = [aij ] \u2208 R N\u00d7N given by aij = 1, if (nj , ni) \u2208 E ; and aij = 0, otherwise. If aij = aji, the digraph is undirected; otherwise is directed. If aij = 1, then j \u2208 Ni. De ne a in-degree matrix D = diag{di} as di = \u2211N j=1 aij . Further de ne a Laplacian matrix L as L = D \u2212 A. Then, de ne a diagonal matrix A0 = diag{ai0} to be a leader adjacency matrix, where ai0 = 1 if and only if the ith vehicle has access to the information of the leader; otherwise ai0 = 0. Finally, de neH = L+A0."
        ]
    },
    {
        "image_filename": "designv11_64_0002084_detc2015-47180-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002084_detc2015-47180-Figure6-1.png",
        "caption": "Figure 6. SCREEN SHOT OF FULL VEHICLE MODEL IN ADAMS [5]",
        "texts": [
            " [15] The vehicle body is modeled as two rigid bodies connected along the roll axis at the center of mass height by a revolute joint and a torsion spring. The influence of body torsion on the vehicle 3 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 08/12/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use dynamics is thus taken into account. The unconstrained degrees of freedom are given in Tab. 1 and a screen shot of the Adams View [5] model is shown in Fig. 6 (Land Rover image from [16]). The simulations were done from the same starting positions as used in the quarter car and pitch-bounce model simulations. Figure 7 shows the vertical acceleration of the center of mass of the full vehicle model with the 4S4 suspension on the \u201cride comfort\u201d setting. The experimental investigation was done on the Belgian Paving at Gerotek Test Facilities [9]. The vehicle was instrumented with a differential GPS (DGPS), three gyroscopes oriented orthogonally and three tri-axial accelerometers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001513_cjme.2015.0302.022-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001513_cjme.2015.0302.022-Figure2-1.png",
        "caption": "Fig. 2. In the present study, the elliptical contact area was divided into numerous narrow bands dy (see in Fig. 2), and the corresponding contact width is 2x, where",
        "texts": [
            " It is only at the boundary of the stick region (y=0) that the offsets coincide with the contact boundary owing to the special geometric area of the ellipse as discussed in Eqs. (1) and (2). The elliptical contact area is defined on the x-y plane, where o is the origin and initial contact point, and the x- and y-axis coincide with a and b as discussed earlier in connection with Fig. 1. The y-axis is in the direction of the cone generatrix, and the sphere rolls in the negative direction of the x-axis. The reversed z-axis is perpendicular to the x-y plane and points inwards of the cone in the rectangular coordinate system. As shown in Fig. 2, c and d denote the semi-axes of the elliptical stick region, which has its midpoint at o' and an offset distance of s. Y ZHAO Yanling, et al: Analysis and Numerical Simulation of Rolling Contact between Sphere and Cone \u00b7524\u00b7 With regard to the distribution of the tangential force, POPOV[31] used similar steps and assumptions as those proposed by CARTER[12] to calculate the threedimensional stress state in the rolling contact area of the sphere and plane. The same procedure was adopted in the present study to obtain the distribution of the tangential force in the three-dimensional rolling contact area"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002715_ijmmp.2016.078055-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002715_ijmmp.2016.078055-Figure2-1.png",
        "caption": "Figure 2 (a) Schematic of laser surface melting and (b) the shape and size of the focused laser beam with laser intensity distribution (see online version for colours)",
        "texts": [
            " The sample was fixed to the working table and the laser beam carried through fibre connected to the sixth arm of robot is traversed at a constant laser power of 4 kW with scanning speed comprising of 10 mm/s. Laser melting has been carried out at fixed working distance of 300 mm to have uniformly focused laser intensity distribution throughout the experimentation. The laser melting has been carried out at 80 J/mm2 calculated using equation (1), dependent on process variables laser power (W), spot size (mm) and scanning speed (mm/s). Figure 2 shows the schematic of laser surface melting process. A nitrogen shroud at 2-bar pressure through a specifically designed and fabricated 25 \u00d7 5 mm off-axis nozzle was used to effectively shield during laser processing against atmospheric contamination. Apparently, nitrogen shielding also provide advantage of diffusing nitrogen in the lasermelted layer as anticipated in previously reported studies (Yue et al., 2006). The laser treatment was carried out along the transverse direction of the plate perpendicular to the rolling direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000961_s11071-014-1713-6-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000961_s11071-014-1713-6-Figure1-1.png",
        "caption": "Fig. 1 Ship schematic illustrating the coordinate systems",
        "texts": [
            " (10): \u03b3 > sgn (so) s2 o (\u2211m i=1wi )2 \u2211m i=1w 2 i [ F2 + \u2211m i=1wi ri \u2211m i=1wi ] (11) The present study uses a nonlinear model for an underactuated marine surface vessel as a test bed for assessing the performance of the observer. Only a very brief description of the model is provided in this Section since the details of the model were covered by the authors in Ref. [27]. The formulation accounts for the six rigid body degrees of freedom of the ship representing the surge, sway, heave, roll, pitch and yaw motions (see Fig. 1). It also considers the rudder dynamics and the physical limitations of the propeller and the rudder. Moreover, it accounts for the effects of coriolis and centripetal accelerations, wave excitations, retardation forces, nonlinear restoring forces, linear damping terms, wind and sea-current loads. This model is referred to hereafter as the full order model of the ship. The wave excitation forces are determined based on long-crested waves with a modified Pierson\u2013Moskowitz spectrum [28]. The linear seakeeping scheme has been incorporated to compute the frequency dependent added-mass and wave-damping terms along with the wave excitation forces [28\u201330]",
            " In the under-actuated configuration of the ship, the actuators are limited to the propeller and the rudder. Thus, the controller is designed based on a reducedorder model of the ship, which only accounts for the surge and yaw motions. As a consequence, the implementation of the controller necessitates accurate estimates of the surge speed, u, the angular velocity of the ship, r , and its time integral, \u222b t 0 rd\u03c4 , around the k \u02dc direction. Note that the required state variables must all be expressed with respect to the body-fixed coordinate system {x, y, z} (see Fig. 1). However, typical measurements in a maritime trajectory tracking appli- cation consist of the heading angle and the ship position with respect to an inertial reference frame {X,Y, Z}. The X and Y coordinates of the vessel are obtained using a global positioning system (GPS) and the heading angle,\u03c8 , is determined by utilizing a gyro compass system. Therefore, one has to extract the values of the required state variables from the measured signals. This is done by first using a nonlinear state observer to accurately estimate X,Y, \u03c8 and their time derivatives with respect to the inertial reference frame {X,Y, Z}"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000991_scis-isis.2014.7044740-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000991_scis-isis.2014.7044740-Figure2-1.png",
        "caption": "Fig. 2. Model of the Wheeled Inverted Pendulum Mobile Robot",
        "texts": [
            " Actually some learning algorithm is necessary because a user has his/her own way to operate the porter robot. This paper proposes a simple learning method for the porter robot based on the wheeled inverted pendulum mobile system. There are two set of parameters to be learned. One is the 978-1-4799-5955-6/14/$31.00 \u00a92014 IEEE 117 parameters of the gaussian models for the human intention recognition. The other is the control parameters for the assistance of the luggage transportation. Experimental results show the validity of the proposed learning method for the porter robot. Figure 2 show the concept of the luggage transportation robot . It is based on inverted pendulum mobile controller that changes the control parameters according to the user intention. It stands by itself and stays while the user leaves it alone. It supports the luggage transportation while the user operates physically the robot with the handle. The robot needs to learn appropriate intention recognition and control parameters user by user because each user has his/her own preference on how to operate the robot",
            " The center of the template \u03bct and the variance \u03c3t include the estimated body posture angle based on the outputs of two-directional accelerometer xa, one rate gyroscope xg, wheel encoder output xe, and input to the wheel motor xu at time t. xt indicates the sensory outputs and motor inputs at time t. The similarity g of the template (\u03bct,\u03c3t) and the sequence of sensory outputs and motor inputs of the robot xt = (xa, xg, xe, ) is calculated as below: g = T\u2211 t=1 exp(\u2212(xt \u2212 \u03bct) T\u03a3t(xt \u2212 \u03bct)) (1) where the T and \u03a3t is the length of the template sequence and the variance matrix that has the element of the variance vector \u03c3t = (\u03c3a, \u03c3g, \u03c3e) in diagonal. T indicates transposition of the vector. Fig. 2 shows the model of our wheeled inverted pendulum mobile robot. The posture angle \u03b8 is estimated by a Kalman 978-1-4799-5955-6/14/$31.00 \u00a92014 IEEE 118 filter using two-directional accelerometer and one rate gyroscope. The Kalman filter is a typical one so that the detailed description of the Kalman filter is omitted in this paper. The wheel angular velocity is measured with an encoder attached to the motor. The posture controller follows conventional torque control theory. Torque for the wheels u is calculated as follows: u = \u2212k1(\u03b8 \u2212 \u03b8d)\u2212 k2\u03b8\u0307 \u2212 k3\u03d5\u0307\u2212 k4 \u222b t (\u03d5\u0307\u2212 \u03d5\u0307d), (2) where \u03b8, \u03b8d, \u03b8\u0307, \u03d5\u0307, \u03d5\u0307d, and t are body angle, desired body angle, angular velocity of the body, wheel angular velocity, desired wheel angular velocity, and time, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000912_iciea.2014.6931429-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000912_iciea.2014.6931429-Figure2-1.png",
        "caption": "Fig. 2 Joint model considering elasticity effect",
        "texts": [
            " ( )T =i ix iy iz\u03d5 \u03d5 \u03d5 and ( )T i ix iy iz\u03b5 \u03b5 \u03b5= stand for the torsional deformation vector and torsional displacement vector, respectively. 1634 2014 IEEE 9th Conference on Industrial Electronics and Applications (ICIEA) The elasticity of joints mainly comes from the servo motor and the harmonic reducer. The gears of the reducer will be equal to torsion spring when the deformation is very small. The manipulator joint model without considering the gap of gears and the transmission error can be shown as Fig.2. In Fig.2, miJ is the moment of inertia of the motor rotor at the joint i , iN is the reduction ratio of the joint reducer, iK is the stiffness coefficient of the torsion spring, i\u03b8 is the rotation angle of the reduced motor, iq is the rotation angle delivered to manipulator link. It is obviously that i\u03b8 is not equals to iq because of the elasticity in joints. In order to indicate the influence on the joint angle caused by the elasticity in joints, a new generalized tinny variable i\u03be is introduced. Then the transformation matrix iB caused by joints\u2019 elasticity can expressed as: 1 0 0 1 0 0 0 0 1 0 0 0 0 1 i i i \u03be \u03be \u2212 =B (6) When establishing the transformation matrix between the two adjacent coordinate systems, the error matrix caused by elasticity of the manipulator links and joints should be considered"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002957_gt2016-56282-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002957_gt2016-56282-Figure3-1.png",
        "caption": "Fig. 3 Photograph of microturbomachinery test rig setup for base excitation in axial direction. Reproduction from Ref. [14].",
        "texts": [
            " The material of foils was Inconel X-750, which has high temperature resistance, and the surface of the top foil was coated by PTFE with a thickness of 20 \u00b5m. The nominal clearance of the test GFTB was adjusted to 155, 180, 205, and 230 \u00b5m by inserting thin metal shims between the bottom foil and the bearing back plate during the extensive dynamic tests. 2 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89521/ on 02/14/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Figure 3 shows a photograph of the microturbomachinery test rig setup with an electromagnetic shaker for base excitation in the rotor axial direction (X-axis). An accelerometer and eddy current sensor measured the axial acceleration of the test rig and the axial motion of the rotor relative to the test rig, respectively, while a shaking force was imposed on the test bench (table). Figure 4 shows a schematic view of the microturbomachinery test rig rigidly mounted on the test bench using a bracket. The rotor had a length and mass of 170 mm and 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003062_s1068366616050044-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003062_s1068366616050044-Figure3-1.png",
        "caption": "Fig. 3. Flow diagram of measurements.",
        "texts": [
            " EXFS from ElvaX allowed the detection of elements from sodium (atomic number is 11) to uranium (atomic number is 92) by recording the spectra at the emitter voltage of 40\u201349 kV. For the exact evaluation of a composition that does not contain light elements, 10 s were sufficient. The resolution is up to 200 eV by the line of 5.9 keV (Fe isotope). The 4096-Channel analog-to-digit transducer was used. CSM tribometer (Fig. 2) was used to determine the volume wear and coefficient of friction in online mode at various temperatures, contact pressures, speed, and moisture content in the lubricating medium. The measurement principle is given in Fig. 3; spherical indenter (sphere) was dropped to the studied counterbody at known precise load. Counterbody represented by disc was mounted on an elastic lever connected to the friction force sensor. During the rotation of the disc sample between counterbody and indenter, the friction force arose, which was measured by minimum deviation of elastic lever using LVDT sensor (inductive sensor of linear displacements). The wear coefficients of the sphere and disc were calculated from the volume of worn material upon the friction of these solids within particular time"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001001_20140824-6-za-1003.00348-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001001_20140824-6-za-1003.00348-Figure1-1.png",
        "caption": "Fig. 1. Overview of DSD forceps",
        "texts": [
            " By compensating for the elongation of the wire using the estimated amount of elongation of the wire, a track ability of the forceps can be improved. In order to demonstrate the validity of the proposed new algorithm, simulation and experimental works were carried out for a simplified wire actuated forceps model. The results showed the effectiveness of the proposed algorithm. The robotic forceps incorporating the screwdrive mechanism, termed double-screw-drive (DSD) mechanism, (DSD forceps) is shown in Fig. 1. The DSD mechanism has three linkages. Two linkages, each consisting of a universal joint of the screwdrive and a spline nut, are for achieving omnidirectional bending motion. A third linkage is for achieving rotary motion of the gripper. 978-3-902823-62-5/2014 \u00a9 IFAC 7233 Opening and closing motions of the gripper are attained by wire actuation. Only one side of the jaws can move, and the other side is fixed. The wire for actuation connects to the drive unit through the inside of the DSD mechanism and the rod, and is pulled by the motor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000804_1.4030097-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000804_1.4030097-Figure2-1.png",
        "caption": "Fig. 2 Connecting rod and bearing geometry",
        "texts": [
            " Both models were meshed using a ten-node tetrahedral mesh apart from the bearing solid where a 27-node hexahedral mesh has been applied. To reduce the computation time only symmetric half models were used. A tied model has been obtained to coincide the bearing and the housing elements. The areas highlighted in red correspond to the stiffness matrix extracted for the bearing lubrication face nodes and the crank journal face nodes. Thereafter, by inverting the stiffness matrices compliance matrices can be obtained [8]. For most general cases, the connecting rod model is constrained in \u201cXYZ\u201d at the shank cut plane as shown in Fig. 2. Furthermore, the crank model is constrained in \u201cXY\u201d at the adjacent main journals and \u201cZ\u201d at a single point. Both material data required including Young\u2019s modulus, Poisson\u2019s ratio, and density for generating the compliance matrix. In addition to the stiffness effects from the bearing compliance matrix, a bearing dynamic deformation file (DDEF) was obtained. The DDEF file has been included in order to specify deformations Contributed by the Combustion and Fuels Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001786_vppc.2015.7352971-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001786_vppc.2015.7352971-Figure1-1.png",
        "caption": "Fig. 1. 2D FEA model of target PMSM",
        "texts": [
            " This method uses the stator current spectrum but the rotor is not rotating. Because of locked rotor, the spectrum of stator current is not affected by the other faults. In order to prevent the demagnetization of PM, this test should be conducted at the lower voltage than rated voltage to reduce the input current of stator winding. In addition, the frequency of input voltage for diagnosis is determined 3 times higher than rated operating frequency to minimize the torque production. 978-1-4673-7637-2/15/$31.00 \u00a92015 IEEE Figure 1 shows the used 2D FEA model of interior permanent magnet synchronous motor (IPMSM). The specifications of IPMSM are summarized in Table 1. The stator of IPMSM is 1 slot skewed to reduce the torque ripple. The skewed structure of stator is considered by converting the 2D analysis result as shown in Fig. 2. In the figure 2, L is the stack length of FEA model, \u03b1 is skew angle, n is the number of an equal division of skew angle. Figure 3 shows the irreversible demagnetization analysis model of IPMSM"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001123_1056789514560916-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001123_1056789514560916-Figure2-1.png",
        "caption": "Figure 2. Geometry of cracked surface.",
        "texts": [
            " Two transverse surface cracks are presented at two locations L1 and L2 with depth b1 and b2 from the fixed end (Figure 3). at Stockholm University Library on August 26, 2015ijd.sagepub.comDownloaded from Both the cracks result in a coupling effect yielding both longitudinal and transverse motion of the shaft. An axial force F1 and bending force F2 are applied at free end of the shaft. Due to the presence of cracks, a local flexibility will be interposed with the order of 2 2 matrix. The geometry of cracked surface is shown in Figure 2. The rate of strain energy released during fracture is Je where Je \u00bc 1 E0 C11 \u00fe C12\u00f0 \u00de 2 \u00f01\u00de For plane strain condition, 1 E0 \u00bc 1 v2 E For plane stress condition, 1 E0 \u00bc 1 E here, g is the Poisson\u2019s ratio and E is the Young\u2019s modulus of elasticity. C11 and C12 represents stress intensity factors of first mode for given load F1 and F2, respectively. The values of local stiffness at Stockholm University Library on August 26, 2015ijd.sagepub.comDownloaded from factor from fracture mechanics can be calculated by taking the rectangular strip d as given in Figure 2. Height of the element w \u00bc 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0R2 2\u00de p Depth of crack h \u00bc w2 \u00f0R b\u00de C11 \u00bc F1 R2 ffiffiffiffiffi h p P1 h w , C12 \u00bc 2F2 R4 w ffiffiffiffiffi h p P2 h w \u00f02\u00de The expressions P1(h/w) and P2(h/w) are the experimental determined functions. The stress intensity factors C11 and C12 can be calculated using two experimental determined functions. So, the following two expressions can be represented as given below P1 h w \u00bc 2w h tan h 2w 0:5 0:752\u00fe 2:02\u00f0h=w\u00de \u00fe 0:37 1 sin\u00f0 h=2w\u00f0 \u00de 3 cos h=2w\u00f0 \u00de \u00f03\u00de P2 h w \u00bc 2w h tan w 2h 0:5 0:923\u00fe 0:199 1 sin\u00f0 h=2w\u00f0 \u00de 4 cos h=2w\u00f0 \u00de \u00f04\u00de Let Vt represent total strain energy due to fracture"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001775_j.procs.2015.12.297-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001775_j.procs.2015.12.297-Figure1-1.png",
        "caption": "Fig. 1. (a) Construction of the flexible cylinder; (b) Construction of the spherical actuator.",
        "texts": [
            " To realize a precise position control of spherical actuator, the flexible displacement sensor is required. For the target, the string type displacement sensor was proposed and tested. However, the frictional force between the slide electrode and the nylon string coated with carbon in the sensor was large, and caused the noise on output signal10, 11. Therefore, it is necessary to develop the flexible sensor with less frictional force and noise. In this paper, a flexible displacement measuring system using wire-type linear encoder is described. Fig.1 (a) shows the construction of the flexible pneumatic cylinder4. The cylinder consists of a flexible tube as a cylinder and gasket, one steel ball as a cylinder head and a slide stage that can move along the outside of the cylinder tube. The steel ball in the tube is pinched by two pairs of brass rollers from both sides of the ball. The operating principle of the cylinder is as follows. When the supply pressure is applied to one side of the cylinder, the inner steel ball is pushed. At the same time, the steel ball pushes the brass rollers and then the slide stage moves toward the opposite side of the pressurized while it deforms the tube. Fig.1 (b) shows the construction of the spherical actuator using flexible pneumatic cylinders. The actuator consists of two ring-shaped flexible pneumatic cylinders which are intersected at right angle and each slide stage of the flexible cylinder is fixed on each handling stage. The size of the actuator is 260 mm in width and 270 mm in height. The total mass of the device is 310 g. In order to measure the attitude angle of each handling stage, two accelerometers are used as angular sensors. Fig.2 shows the transient view and the response of the stage angle of the spherical actuator"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001201_amr.1017.624-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001201_amr.1017.624-Figure3-1.png",
        "caption": "Fig. 3 Schematic diagram of wireless measurement tool holder system",
        "texts": [
            " This figure shows that the monitoring method effectively estimated the end-mill process by using a high response infrared thermograph that can continuously transmit frame images to a personal computer. Workpiece End-mill Themography \u03b8 \u03b8 = 150\u00b0 \u03c9 1.5 m Top view Tool holder End-mill tool Workpiece Themography Tool holder Workpiece Machined chips Reflected image Machined surface End-mill tool Workpiece: JIS SUS310S Cutting tool: OSG WXL-EMS D = 10 mm, 4 flute Machining conditions: Vc = 45 m/min, fz = 0.05 mm/tooth, Ad = 12 mm, Rd = 0.6 mm, dry air Fig. 1 Setup for monitoring end-mill process Fig. 2 Infrared imagery of end-mill process Tool for measuring internal temperature. Figure 3 shows the schematic diagram of the wireless measurement tool holder system. If the measurement object was a rotating tool, there was the problem that the wire wound around the spindle when it rotated at high speed. Therefore, we have developed a wireless measurement tool holder system. First, we inserted the thermocouple into the hole in line with the tool center axis . Next, we placed the amplifier, A/D converter, micro-controller, and transmitter in the tool holder. The tool temperature measurement results were wirelessly and continuously transmitted to a computer that is connected to the receiver"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003134_978-981-10-2875-5_70-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003134_978-981-10-2875-5_70-Figure9-1.png",
        "caption": "Fig. 9 Kinematic sketch of the tetrahedral element",
        "texts": [
            " Based on the structural topology relationship of the minimum composite unit, we can derive that the number of the DOFs of the whole deployable truss antenna shown in Fig. 1 also is one. That is to say, only the positions of the connection nodes of the deployable truss antenna are changing during its deploying/folding process, but their orientations keep constant. It can be known from the analysis mentioned in Sect. 2 that the orientations of the connection nodes are unchangeable during the deploying/folding process of the antenna. Consequently, the kinematic sketch shown in Fig. 9 is used to analyze the position and velocity variations of the nodes of the tetrahedral element. As shown in Fig. 9, a reference coordinate system H-xyz is attached to the node H with x-axis parallel to the BA and z-axis along the OH. Then the following equation can be obtained: BP \u00bc l1 sin u; BO \u00bc AO \u00bc CO \u00bc l1 cos a1 sin u; PO \u00bc l1 tan a1 sin u; \u00f07\u00de where, cos a1 \u00bc l1 sin a1 \u00fe a2\u00f0 \u00de=l2, cos a1 \u00fe a2\u00f0 \u00de \u00bc l21 \u00fe l23 l22 = 2l1l3\u00f0 \u00de, l1, l2 and l3 represent the length of the struts AM (BM), AN (CN) and BK (CK), respectively, u denotes half of the angle between the BM and AM, a1 (a2) is the constant angle between the BO and BA (BC)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002683_978-3-319-11930-4-Figure4.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002683_978-3-319-11930-4-Figure4.1-1.png",
        "caption": "Fig. 4.1 Cross section of ball\u2013rod\u2013race loading and thin-film locations",
        "texts": [
            " Eventually, coating spall will occur at the contact surface which further increases shear loading beneath the surface of the ball and furthers the onset of crack initiation below the surface. Subsurface cracking and spalling may also result from internal material defects. The presence material defects in the lattice structure of the contacting materials of both the rolling elements and the lubricating film is a large contributor to the wide scatter in RCF test results. The RCFmethod in this chapter uses ball bearings and a rotating rod to load thinfilm coatings in high-cycle fatigue. Coatings on either the balls or the rotating rod may be tested as shown in Fig. 4.1. The RCF test method was first conceived in 1958 by W. J. Anderson at NASA Lewis for gear and bearing research. Originally named the \u201cNASA 5-ball fatigue tester,\u201d Anderson constructed the tester in 1959, and today it is simply known as the ball\u2013rod\u2013cup RCF tester. The ASTM publication STP771 recommends \u00bd\u201d diameter balls made from M50 tool steel be used for consistent testing of candidate rod materials. The ball\u2013rod\u2013cup test configuration has matured such that the balls and cups may be purchased from leading bearing component manufacturers",
            "31E-10 Reproduced with permission from Wear, Volumes 274\u2013275, 27 January 2012, Pages 368\u2013376, Elsevier 82 4 Rolling Contact Fatigue in High Vacuum L10 \u00bc C W 3 ; \u00f04:17\u00de where the variableW corresponds to the radial load applied to the ball and C is the basic load capacity of the test configuration with respect to a ball bearing-type system. The basic load capacity C may be calculated using the RCF test results presented in Figs. 4.8, 4.9, and 4.10. The values L10 and W are calculated and defined for each test configuration from Fig. 4.1. Figure 4.21 presents basic load capacity calculations for the test configurations presented in Table 4.1. Also plotted in Fig. 4.21 are the stress cycle values based on the L10 life from the RCF experiments of Figs. 4.8, 4.9, and 4.10, represented as vertical lines of a specific type (dashed, hashed, dotted). It is interesting that configurations 1 and 3 give different load capacities for the same configurations, suggesting that life-adjustment factors are needed to accurately calculate L10 life for these test element combinations",
            " The rolling contact fatigue test platform presented in Chap. 4 was used to quantify the RCF life of the extreme DoE coated balls. The tests were conducted in high vacuum in the range of 10 7 Torr using a fixed load and speed for all tests. The Hertz contact stress was calculated as 4.1 GPa with a rotation speed of 130 Hz (7,800 rpm). For clarity, Fig. 6.6 contains a cross section of the cup and ball configuration and the proximity of the balls and rod for the RCF setup. This setup is the same as Fig. 4.1 in Chap. 4, except that the balls are coated with ion-plated nickel\u2013copper\u2013silver instead of evaporated pure silver. In addition to the validation tests of Chap. 4, the RCF test platform was further validated for repeatability using multielement coatings on 7.94 mm balls for nine tests. These tests used a separate set of nickel\u2013copper\u2013silver-coated balls, all from the same lot and process history. The coated balls were purchased from an outside ball coating supplier and represent an outside coating system that is independent of the coatings presented in Chap",
            " In this chapter a real-time target application is used to test and characterize control schemes related to argon pressure control, plasma total-current monitoring, and argon process gas regulation to maintain optimum plasma properties during deposition. The primary benefit of HIL testing is controller development and optimization using real system hardware instead of models or prototypes. In addition, a second benefit is safe operation of the equipment components. The HIL process for controller development allows safe limits to apply to the hardware motion to prevent damage due to incorrectly tuned equipment. All HIL testing in this chapter was carried out using the ion-plating system presented in Fig. 4.1 in Chap. 4. The ion-plating process requires multiple control systems to operate independent of each other at the subsystem level, but concurrently as a process. The Simulink Real-Time tool requires an external computing platform, the xPC host machine, from which the model will be executed. The xPC host allows interface with all component hardware through either analog or digital \u00a9 Springer International Publishing Switzerland 2015 M. Danyluk, A. Dhingra, Rolling Contact Fatigue in a Vacuum, DOI 10",
            " Similar plant model information was gathered for the input gas plant model Q1. All first-order plant models were calculated as described in Eqs. 8.1, 8.2, and 8.3. From the RCF life test results of Chaps. 6 and 7, the process pressure corresponding to the longest RCF life is between 15 and 17.5 mTorr. The inlet and exit flows, Q1 and Q2, were set to provide net flow of 60 sccm, which corresponds to approximately 15.5 mTorr steady-state pressure inside the chamber using the cryogenic pump system on the ion plater of Fig. 4.1. 152 8 Real-Time Process Control Referring to Fig. 8.4, the flow of the step input disturbance test ranged from 60 to 82 to 60 sccm over time. During the test, the flowsQ1 andQ2 were held constant to maintain 15.5 mTorr pressure and only the disturbance input,Qd, was increased for 2 s to 22 sccm. There is a 2 s delay in the pressure monitor response as well as an overshoot after the disturbance Qd is switched off. To summarize, the pressure monitor system was late to respond to the disturbance, which led to a delayed response or the conductance valve to maintain requested chamber pressure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003730_0954406213519616-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003730_0954406213519616-Figure5-1.png",
        "caption": "Figure 5. (a) Example 1: System under consideration, and (b) angles and unit vectors defining positions of the springs and dampers.",
        "texts": [
            " The sum of the damping coefficients of the dampers from the equivalent systems is given by cI \u00fe cII \u00bc XN j\u00bc1 cj \u00f026\u00de Proof. The relationship (26) is easily obtained by summing up equation (21a,b). In order to illustrate the use and benefits of the approach presented earlier, two examples are discussed subsequently. Example 1. The system under consideration consists of three linear springs and three linear viscous dampers whose positions as well as stiffness and damping coefficients are shown next to them in Figure 5(a). All the elements are connected in the centre, where the particle of mass m is located. Let us find the equivalent system of two mutually orthogonal springs and two mutually orthogonal dampers. In addition, we will also show how this approach easily gives the natural frequencies of free undamped vibration. Following the notation from Figure 1(a) and using explanations given in Figure 5(b), we identify the angles between the axis of the springs and dampers as \u20191 \u00bc 4 =3, \u20192 \u00bc 2 =3, \u20193 \u00bc 0, \u20191 \u00bc 5 =3, \u20192 \u00bc =3, \u20193 \u00bc . The two-element system of dampers equivalent to the one shown in Figure 5(a) is obtained by calculating first the position of one of the dampers, i.e. the angle c (see equation (20)) tan 2 c \u00bc P3 j\u00bc1 cj sin 2 \u2019jP3 j\u00bc1 cj cos 2 \u2019j \u00bc 2c sin 10 =3\u00f0 \u00de \u00fe c sin 2 =3\u00f0 \u00de \u00fe 3c sin 2 \u00f0 \u00de 2c cos 10 =3\u00f0 \u00de \u00fe c cos 2 =3\u00f0 \u00de \u00fe 3c cos 2 \u00f0 \u00de \u00bc ffiffiffi 3 p 3 \u00f027\u00de the solution of which is c \u00bc 12 \u00bc 15o \u00f028\u00de Equation (21a,b) yields the values of two damping coefficients cI \u00bc X3 j\u00bc1 cj cos 2 \u2019j c \u00bc 2c cos2 5 =3\u00fe =12\u00f0 \u00de \u00fe c cos2 =3\u00fe =12\u00f0 \u00de \u00fe 3c cos2 \u00fe =12\u00f0 \u00de \u00bc 3:866 c, cII \u00bc X3 j\u00bc1 cj sin 2 \u2019j c \u00bc 2c sin2 5 =3\u00fe =12\u00f0 \u00de \u00fe c sin2 =3\u00fe =12\u00f0 \u00de \u00fe 3c sin2 \u00fe =12\u00f0 \u00de \u00bc 2:134 c \u00f029a; b\u00de These two dampers are shown in Figure 6",
            " If there is no damping, the natural frequencies of free vibration are easily calculated as !1 \u00bc ffiffiffiffiffiffi kII m r 1:461 ffiffiffiffi k m r , !2 \u00bc ffiffiffiffi kI m r 1:966 ffiffiffiffi k m r \u00f033a; b\u00de at NORTH CAROLINA STATE UNIV on May 11, 2015pic.sagepub.comDownloaded from In order to emphasize some of the benefits of the approach presented earlier, let us show the alternative way of finding the natural frequencies. The springs are undeformed (unstretched) in the stable static equilibrium position around which the particle performs small oscillations (Figure 5) and, consequently, one needs to determine the deflection in the direction that is collinear with the direction of the spring in the static equilibrium position (see Appendix 1 for detailed explanations of this fact). Based on Figure 7(a), these deflections are found to be 1\u00bc x=2\u00fey ffiffiffi 3 p =2 , 2\u00bc x=2 y ffiffiffi 3 p =2 , 3\u00bc x \u00f034a c\u00de and the potential energy is V \u00bc 1 2 k1 2 1 \u00fe 1 2 k2 2 2 \u00fe 1 2 k3 2 3 \u00bc 15k 8 x2 ffiffiffi 3 p k 4 xy\u00fe 9k 8 y2 \u00f035\u00de The kinetic energy of the particle is T \u00bc m _x2 \u00fe _y2 =2, so that Lagrange\u2019s equations of the second kind for small oscillations d dt @T @ _q \u00fe @V @q \u00bc 0, q 2 x, y \u00f036\u00de yield the following differential equation 4m \u20acx\u00fe 15kx ffiffiffi 3 p ky \u00bc 0, 4m \u20acy ffiffiffi 3 p kx\u00fe 9ky \u00bc 0 \u00f037a; b\u00de Assuming the solution for motion in the form x \u00bc A cos "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001972_ilt-03-2015-0034-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001972_ilt-03-2015-0034-Figure4-1.png",
        "caption": "Figure 4 Representation of the studied geometry",
        "texts": [
            " The numerical method is detailed in the document ANSYS CFX-Solver Theory (2003); this algorithm makes use of the interpolation method of Rhie and Chow (1982) and to prevent disturbance of the field pressure. This method is among the best methods to save memory space and computing time. Figure 3 shows the general flowchart of the code ANSYS-CFX calculated. The shaft made of steel is driven by an electric motor of 21 kW with a diameter of 99.82 mm, a bushing of an internal diameter of 100 mm and an external diameter of 140 mm, (Figure 4). This results in a bearing with a radial clearance of 90 m. The first bearing layer is a 38-mm-thick steel structure, whereas the second layer-internal is a low-friction material (tin) with a thickness of 2 mm (88 per cent Babbitt ASTM B23-949). The loading system relies on a pneumatic cylinder that acts on the bearing sleeve. The load is applied to the bearing housing along the vertical direction without friction using a static hydraulic pressure with the same fluid as the tested oil. This geometric model is a real case from an experimental study and numerical analyses on elastic effect in journal bearing severs loading published (2012, 2014)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003408_imece2016-67407-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003408_imece2016-67407-Figure7-1.png",
        "caption": "Figure 7. Square rod under uniaxial tension in z-axis P",
        "texts": [
            " In addition, the emergence of size effects depends on the ratio of the free edge to the solid cross sectional area. In order to include the size effects in the stiffness tensor (Equation 7), the effective elastic modulus Ef should replace the constitutive material elastic modulus Es. The free perimeter to solid cross sectional area ratio also results in an anisotropic characteristic as evident from the following case. Consider a square rod of side a and length L, under two different loading cases. The first case is applying a uniaxial tensile load P along zaxis as shown in Figure 7. The effective elastic modulus is related to the bulk and surface moduli as follows: 6 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90996/ on 02/15/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use \ud835\udc38\ud835\udc53 = \ud835\udc38\ud835\udc4f + 4\ud835\udc51 \ud835\udc4e \ud835\udc38\ud835\udc60 (21) The second case is applying a uniaxial tensile load P along y-axis. The elastic moduli relation would be as follows: \ud835\udc38\ud835\udc53 = \ud835\udc38\ud835\udc4f + 2(\ud835\udc4e + \ud835\udc3f) \ud835\udc51 \ud835\udc4e \ud835\udc3f \ud835\udc38\ud835\udc60 (22) By comparing Equations (20) and (21), we can conclude that there would be a different effective elastic modulus in different loading directions depending on the free perimeter to cross sectional area ratio perpendicular to the loading direction, which represents an anisotropic property of the \u201cGurtin Murdoch\u201d surface elasticity model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000592_9781118899076.ch8-Figure8.19-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000592_9781118899076.ch8-Figure8.19-1.png",
        "caption": "FIGURE 8.19 Orientation of electron spins in the spin-frustrated ground state of NI (a) and the calculated geometric structure of NI (b). Reproduced with permission from Ref. [34]. \u00a9 2011 RSC Publishing.",
        "texts": [
            " Importantly, the MCD intensity at high field does not decrease in intensity with varying temperature as 1/T, the Curie law behavior predicted for an isolated S =\u00bd ground state (Fig. 8.18d). For instance, the intensity at 25,000 cm\u22121 first decreases and then increases in intensity (Fig. 8.18e), consistent with the Boltzmann population of a low-lying excited state at approximately 150 cm\u22121 above the S =\u00bd ground state and with a different MCD signal (Fig. 8.18f ). This low-lying excited state requires all three coppers of the TNC are close to equally AF coupled (~500 cm\u22121). This situation results in a spin-frustrated ground state for NI. Figure 8.19a shows that AF coupling of Cu1 with Cu2 and Cu2 with Cu3 leads to a parallel alignment of the spins on Cu3 and Cu1, although theywant to be AF coupled due to the superexchange pathway associated with a Cu1\u2013Cu3 bridging ligand. Therefore, all three coppers must be bridged by the products of complete dioxygen reduction. Additionally, the unique g < 2.0 values, which originate from a phenomenon known as antisymmetric exchange that is associated with this spin frustration [52], and the sign and nature of the temperature dependence of the pseudo-A term in the MCD spectra in (Fig. 8.18a, bottom) unambiguously lead to the structural assignment of NI as containing a \u03bc3-oxo ligand bridging all three of the coppers of the TNC and a \u03bc2-OH between the T3s (Fig. 8.19b) where the two bridging oxygen ligands originate from dioxygen reduction [53, 54]. Having defined the electronic and geometric structures of PI and NI, the mechanism of the 2nd two-electron reductive cleavage of the O\u2500O bond was explored [49]. 184 MOLECULAR PROPERTIES AND REACTION MECHANISM (a) 6000 SF-Abs RFO-LT-MCD 10 msec 1 msec 365 nm 318 nm T1 T1 7T ^ 1.8K at 27,560 cm\u20131 31,780 cm\u20131 N1 N1 4000 2000 40 20 0 \u201320 \u201340 100 50 80 60 40 20 0 0 50 100 150 Temperature (K) Temperature dependence of MCD intensityX-band EPR 25,000 cm\u20131 1000 250 150 100 70 AE (cm\u20131)0 \u201350 \u2013100 15,000 8970 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002464_978-981-10-1109-2_5-Figure5.3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002464_978-981-10-1109-2_5-Figure5.3-1.png",
        "caption": "Fig. 5.3 1 Isometric view of a partially folded fPZM. 2 Section A\u2013A with \u03c8 , \u03b6 and \u03b6 \u2217 in the same plane. \u03c8 and \u03b6 are highlighted in: red and green, respectively. 3 Plot showing the relationship between angles \u03c8 and \u03b6 \u2217",
        "texts": [
            " Preliminary experimentation with the physical models including the one shown in Fig. 5.1 suggested that fPZM is a proper mechanism. Further geometrical analysis has also shown that, as illustrated in Fig. 5.2. Although the fold is a function of the side angles \u03c8 , it is linked to the angle (\u03b6 \u2217) between faces T and B of a fPZM. Obviously, for \u03c8 equal to 0 and \u03c0 , the values of this corresponding central angle are: 0 and \u03b6 , respectively. Angle \u03c8 does not depend on the number of sides n. The trigonometrical relationships of a folded triangular fPZM (n = 3) are shown in Fig. 5.3. Based on the relationships shown in Fig. 5.3, the intermediate values of the central angle \u03b6 \u2217 for fPZM during folding are calculated as follows: \u03b6 \u2217 = 2 arcsin [ sin \u03b6 2 \u00d7 sin \u03c8 2 ] (5.1) where sin( \u03b6 2 ) is a constant parameter for a given fPZM. Since the module is formed from rigid plates with revolute hinges only, such folding seems practical and intuitive. This is expected to facilitate the processes of: deployment & stowing and potential sealing & pressurization. Similar folding concept has been considered for a human lunar base in [7]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.7-1.png",
        "caption": "Fig. 10.7 Configuration of drilling",
        "texts": [
            " For biomedical devices, there is sometimes an urgent need for producing a custom-made device in a necessary time frame, and therefore, process planning is particularly important in biomedical/clinical applications. In order to determine the most feasible machine operation parameters to achieve a material removal process with minimal manufacturing cost and time, we should understand the relations between the different operation parameters and the manufacturing concerns, such as machine time and machine removal rate (MRR). The typical configuration of drilling is shown in Fig. 10.7. The drilling head with a diameter of D rotates at a machine speed (N ) with a unit of revolution per minute (rpm), and therefore, the cutting speed v (unit: mm per min) at the tip of cutting edge is v\u00bc \u03c0ND. In other words, if a process requires a cutting speed v, we need to set the drilling speed accordingly. In practice, we might wish to apply a higher drilling speed such that an ideal process is guaranteed: N \u00bc kv \u03c0D \u00f010:1\u00de where k is the unit constant, whose suggested value can be checked from a manufacturing handbook"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000318_tia.2010.2070784-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000318_tia.2010.2070784-Figure6-1.png",
        "caption": "Fig. 6. Three examples of the remaining phase coils after removing one of the two three-phase DL windings. (a) DL-1. (b) DL-2. (c) DL-3.",
        "texts": [
            " Therefore, the derived SL winding results to have a number of turns Nt that are half of the allowable rated number. The inverter that is available for the tests has six separate fullbridge converters. It is used for both three- and six-phase supplies, and its maximum phase current is equal to 6.2 A (peak). In this section, some considerations are carried out about the arrangement of the two sets of three-phase windings. Both healthy and faulty operating conditions are taken into consideration. The DL windings are shown in Fig. 6, in which only one of the two three-phase windings is shown. In configuration DL-1, each of the two three-phase windings is placed in a well-defined part of the stator [Fig. 6(a)]. On the contrary, the coils of the two windings are alternated along all the stator circumference in configurations DL-2 [Fig. 6(b)] and DL-3 [Fig. 6(c)]. Similarly, the SL windings are shown in Fig. 7. In the SL winding, the coils are wound around every other tooth, and each coil should be formed by a double number of turns (however, it is not possible in the prototype considered here). Configuration SL-1 has the two windings concentrated in a part of the stator [Fig. 7(b)]. Configuration SL-2 has the two windings with alternated coils [Fig. 7(c)]. 1) Healthy Operating Conditions: Depending on the winding arrangements, it is possible to take advantage of the presence of the two separate three-phase windings",
            " Six different angular positions of the phase coils are required, according to the six angular positions of the phase current vectors. A mechanical shift is necessary between phases A, B, and C and phases A\u2032, B\u2032, and C \u2032, corresponding to 30 electrical degrees, with the motor resulting to be exactly a six-phase motor. In the DL configurations, the six-phase supply is possible when the ratio Q/2t is even, where Q is the number of slots and t is the machine periodicity [10]. In the 12-slot 10-pole motor, a six-phase supply can be used only when the DL-3 configuration is adopted [see Fig. 6(c)]. 2) Faulty Operating Conditions: Operations under faulty conditions have to be considered as well during the winding design. A simple split of the stator in two parts is not enough, as will be shown in the next section. When one winding is removed and only the remaining winding is supplied, the performance of the PM motor changes considerably according to the adopted coil distribution. The main effect of an erroneous arrangement can be an unbalanced force on the stator and rotor (and on the bearings)",
            " Summarizing, with one winding supplied, the average torque is almost halved, independent of the winding arrangement and rotor topology. On the other hand, there is a general increase of the torque ripple in faulty conditions, as confirmed by measurements. It is worth considering the peak of the unbalanced radial force on the rotor, reported in the aforementioned tables. The worst case is achieved by adopting configuration DL-1 regardless of the motor topology. This is unsurprising since the winding covers only a part of the stator [Fig. 6(a)]. A nonnegligible radial force is also observed when configuration DL-2 is adopted, as well as the SL windings (configurations SL-1 and SL-2). In these cases, the motor does not remain symmetric. Conversely, the DL-3 configuration is characterized by coils of the same phase wound on the opposite part of the stator. Thus, the pull forces are always balanced, achieving a very low unbalanced radial force. The unbalanced radial forces found using the SPM motor are prominently lower than the forces using the IPM motor",
            " Finite-element simulations are carried out, imposing constant currents in the phases of one winding (e.g., phases A, B, and C) and considering the other winding disconnected. In particular, IA = I\u0302 = 1 A, and IB = IC = \u2212I\u0302/2 = \u22120.5 A. The PM is removed from the rotor so as to have a zero PM flux. In addition, the IPM rotor iron bridges (normally saturated by the PM flux) are removed as well. Adopting configuration DL-1, there is no interaction between the two three-phase windings. This is reasonable since the two windings are placed in two separate parts of the stator [see Fig. 6(a)]. Moreover, configuration DL-2 shows a low interaction between the windings, with the coils of each phase being grouped. However, such a decoupling does not occur with the DL-3 configuration, showing the higher mutual coupling between the two windings. This is a direct consequence of the arrangement of the coils of the two windings, which are alternated along the stator circumference, as shown in Fig. 6(c). The flux induced by the phase coils, wound on the opposite teeth, links mainly the two adjacent coils (that refer to two different phases of the other winding). The flux induced in the disconnected phases approaches half the value of the flux linkage of the supplied phases. Finally, both configurations adopting the SL winding yield a low mutual coupling between the two windings. Some tests were carried out on the IPM motor prototype. The three phases of one winding (e.g., A, B, and C) are supplied by means of alternate voltages at 50 Hz, while the phases of the other winding (e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002790_jera.25.1-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002790_jera.25.1-Figure1-1.png",
        "caption": "Figure 1 The circuit of the MAF process applied in research: \u0430) - side view; b) - view from above; 1 - clamp, 2 \u2013blank or a sample, 3 - magnetic pieces, 4 \u2013 magnetic abrasive powder",
        "texts": [
            " In articles [3-5] the authors have applied MAF method to finish edges on cutting tools to increase of their durability. The authors of work [6] have shown an opportunity to use MAF for removal micro burrs and creation precision finishing edges on electronic devices components. In given article the results of research directed to raise productivity of deburring for not rigid blanks by MAF method are reflected. Optimization of magnetic abrasive deburring conditions Experiment conditions. The circuit of the MAF process chosen for experiments is shown in Fig. 1. The blank 2 fixed by a clip 1 is moved between magnetic pieces 3 on an arc trajectory with the district speed vrot and oscillates in a vertical plane with an average speed vosc. At the movement the blank overcomes resistance of a magnetic abrasive powder (further \u2013 powder) filling up the work zone kmpr. The magnetic abrasive powder is kept in the work zone by magnetic forces and by friction in the contacts of the powder with the magnetic pieces. Rigidity of the abrasive media in the work zone and intensity of its influence on blanks was changed by varying a magnetic induction on the magnetic pieces 3 of a magneto generator 1 ",
            " For this reason we used a specific index of the allowance removal q1, mg/(cm 2 \u00b7s) to estimate results of the experiments at optimization of MAF conditions, - a metal weight, removed from 1 cm 2 of the work surface at one second of a blank/sample presence in the work zone. Burrs on sample edges were previously deleted in order to diminish received meanings q1 dispersion at the stage of the conditions MAF optimization. For first experiments work surfaces of the magnetic pieces were carried out with the cylindrical form and with identical radiuses Re = Ri =150 mm (see Fig. 1). Because of this the work zone had the crescent form in the form of above. The width \u03b4 of the work zone decreases in process of moving away from an average part to areas km and pr. Edges of the magnetic pieces limiting the work zone kmpr are concentrators of a magnetic flow, and reduction of \u03b4 is accompanied by reduction of magnetic resistance of the powder media. For the named reasons density of the magnetic abrasive media should be more in areas km and pr, than in the average part of the work zone, and it should promote intensification the allowance removal there"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000249_josr.57.1.120007-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000249_josr.57.1.120007-Figure5-1.png",
        "caption": "Fig. 5 CADETand BNTship lines hull geometry",
        "texts": [],
        "surrounding_texts": [
            "Ankudinov and Jakobsen (1996) and Ankudinov et al. (1996, 2000) proposed the MARSIM 2000 formula for maximum squat based on Sm and Tr in shallow water. The Ankudinov method has undergone considerable revision as new data were collected and compared. The most recent modifications from a study of ship squat in the St. Lawrence Seaway by Stocks, Daggett, and Page (2002) and correspondence between Ankudinov and Briggs in April 2009 are contained in this study. These new revisions were programmed and documented in a technical note by Briggs (2009). The Ankudinov formula has been used extensively in the CHL Ship Tow Simulator. The Ankudinov prediction is one of the most complicated formulas for predicting ship squat because it includes many empirical factors to account for the effects of ship and channel. The restriction Fnh 0.6 is applied. The maximum ship squat, SMax, is a function of Sm and Tr given by SMax \u00bc Sb Ss \u00bc Lpp Sm 0:5Tr\u00f0 \u00de Lpp Sm \u00fe 0:5Tr\u00f0 \u00de \u00f03\u00de The SMax can be at the bow or stern depending on the value of Tr. The negative sign is used for bow squat, Sb, and the positive sign for stern squat, Ss."
        ]
    },
    {
        "image_filename": "designv11_64_0002864_0954405416661003-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002864_0954405416661003-Figure7-1.png",
        "caption": "Figure 7. Model of spiroid gear.",
        "texts": [],
        "surrounding_texts": [
            "Two types of geometry of spiroid gears that were manufactured by CNC machining have been provided to validate the generation of surface modifications of the gears by using the ease-off topography (Figures 7 and 8). NC codes are programmed automatically by use of Unigraphics NX software. The mated geometry of spiroid pinion is generated by the hob process. An alloy of Cu + Zn is selected as the materials of gears."
        ]
    },
    {
        "image_filename": "designv11_64_0000508_pssa.201532241-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000508_pssa.201532241-Figure1-1.png",
        "caption": "Figure 1 (a) Chemical structure of 2,4-bis[4-(N,Ndiisobutylamino)-2,6-dihydroxyphenyl] squaraine. (b) Optical micrograph of the squaraine crystal, fixed on the SiO2/Si substrate via a polystyrene film (PS), and the deposited gold electrodes. The perimeter of the PS film is denoted by a dashed white line. In addition, the unit cell of the squaraine crystal is shown with dimensions given in the text. (c) Schematic illustration of the SQ crystal with four gold electrodes shown in (b). The size of the crystal amounts to b0 = 44 m, c0 = 36 m, and a0 = 440 m. Additionally, the electric circuits for the four-probe measurement are shown: V14 and I14 are the bias voltage and the current between electrodes 1 and 4, respectively; V2 and V3 are potentials of the electrodes 2 and 3, respectively.",
        "texts": [
            " We found: (i) the contact resistance of all electrodes was larger than the resistance of the squaraine crystal (Rsq); (ii) for the same electrode, the values of the resistance were different for injecting (Rin) and extracting (Rex) charge carriers; (iii) the values of all eight contact resistances (four injecting and four extracting charge carriers) were not constant, but depended on the current. Furthermore, all derived values of the resistances were found to be within two ranges: (a) Rin > Rex > Rsq and (b) Rex > Rin > Rsq. Each of the measured I\u2013V characteristics was allowed to identify three components, representing charge injection, charge transfer through squaraine crystal, and charge extraction. 2 Experimental Figure 1(a) schematically illustrates the chemical structure of the investigated molecules, 2,4-bis[4-(N,N-diisobutylamino)-2,6-dihydroxyphenyl] squaraine (SQ) (empirical formula C32H44N2O6) purchased www.pss-a.com \u00a9 2015 WILEY-VCH Verlag GmbH & Co. KGaA, Weinheim p h ys ic ap s sst at u s so lid i a from Aldrich. First, a SQ crystal was placed onto the substrate consisting of a 230 nm SiO2 layer on 0.7 mm thick Si support. The crystal was fixed to the substrate by surrounding it with a layer of polystyrene film deposited from 1% cyclohexanone solution [see Fig. 1(b)]. The SQ crystals were typically 30\u201350 m wide, 30\u201350 m thick, and 300\u2013500 m long. In addition, Fig. 1(b) shows the unit cell of the SQ crystal (monoclinic) with dimensions a = 6.2034 A\u030a, b = 16.478 A\u030a, c = 14.518 A\u030a; and \u03b1 = 90\u25e6, \u03b2 = 92.406\u25e6, \u03b3 = 90\u25e6, as determined in Ref. [30]. 100 nm thick Au electrodes were evaporated onto the SQ crystals through a mask which had windows of a width of 30 m. During the deposition process, the sample was tilted, allowing to achieve continuous Au electrodes on the crystal [see Fig. 1(b) and (c)]. As a result, the Au electrodes contacted the faces along the b- and c-axis, and charge transfer was measured along the a-axis, i.e., the \u2212 stacking direction. As a control experiment for testing the sensitivity of our set-up, we have measured the leakage current through a polystyrene film. For voltages up to 20 V, we measured a current below 10 fA (see Supporting Information, online at: www.pss-a.com). Electrical characterization of each sample was carried out by measurement of total twelve I\u2013V characteristics for each crystal, based on a four-, three-, and twoprobe configuration, respectively [see Fig. 1(c)]. A subfemtoAmpere meter source (Keithley 6430) was used to apply the voltage V , and the current I was measured. All I\u2013V characteristics were recorded at an ambient temperature 300 K with the bias voltage V ranging from 0 to 15 V (in 5 mV steps). We denoted the electrode, where the positive pole of the voltage was applied, as the injecting electrode. Correspondingly, the electrode, where the negative pole of the voltage was applied, we denoted as extracting electrode. The latter was grounded",
            "5 V, the voltage drop V14 is distributed over the whole range between electrodes 1 and 4; (iii) for r41, the equality of potentials V4, V3, and V2 means that the voltage drop on the extracting electrode 1 is dominant. According to the results presented in Fig. 2, the SQ crystal with four gold electrodes can be represented as shown in Fig. 3(a). For comparing the I\u2013V curves of four-, three-, and two-probe measurements (and also for the various samples), we have introduced a resistance unit for a SQ crystal, denoted as Rsq, which was determined as the resistance of the crystal having a volume of a0 \u00d7 b0 \u00d7 c0 = 24.2 \u00d7 27.5 \u00d7 26.4 m3 [see denotation in Fig. 1(c)] . Assuming that the cross-section of the layer of certain thickness in the SQ crystal, where charge carrier transport took place, was constant for all four-, three-, and two-probe measurements, the squaraine partial resistances Rii and Rij can be found as nii \u00d7 Rsq. There, the indices i and j vary between 1 and 4. The coefficients are given by nii = Lii/24.2 m and nij = Lij/24.2 m, with Lii and Lij being the corresponding distances as shown in Fig. 3(a). For example, for the sample shown in Fig. 1(b): L11 = 30 m, L12 = 26 m, L22 = 35 m, L23 = 24 m, L33 = 37 m, L34 = 24 m, and L44 = 35 m. The corresponding values of Rij and Rii are: R12 = 1.1 \u00d7 Rsq, R22 = 1.4 \u00d7 Rsq, R23 = Rsq, R33 = 1.5 \u00d7 Rsq, and R34 = Rsq. The relating electric circuits of the four-, three-, and two-probe measurements for both current directions are schematically illustrated in Figs. 3(b)\u2013(g). The resistance r14 of the device between the electrodes 1 and 4 can be represented as r14 = R1in + R12 + R22 + R23 + R33 + R34 + R4ex = R1in + R4ex + 6 \u00d7 Rsq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001184_icphm.2014.7036369-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001184_icphm.2014.7036369-Figure2-1.png",
        "caption": "Figure 2. The illustration of the test bench.",
        "texts": [
            " The biphase component of each BRWB sample has been made independent over each other in previous steps. Thus, by the average operation, the spurious bicoherence coming from long coherence time waves can be eliminated automatically. Commonly, 10-20 signal epochs are sufficient to obtain a reliable BRWB estimation [18]. Step 8: Features estimation by Eq. (6), Eq. (7) and Eq. (8). IV. FAULT DIAGNOSIS FOR LOCOMOTIVE ROLLER BEARINGS Experiments are conducted on the bearing test bench, which is shown in Figure 1. The illustration of this test bench is shown in Figure 2. The outer race of the locomotive bearing is driven by a Nylon driving wheel which is connected to a hydraulic motor, while the inner race of the bearing is keep fixed in the measuring process. In this experiment, four bearing health conditions are considered, which include inner race defect, outer race defect, ball defect and normal bearing. Since vibration-based analysis is one of the principal tools for mechanical fault diagnosis, the vibration is measured by the accelerometers mounted on the bearing house at the motor speed 450r/min"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000802_amm.658.225-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000802_amm.658.225-Figure2-1.png",
        "caption": "Fig. 2. The variation of concentrated force on fingerboard",
        "texts": [
            "159, Pennsylvania State University, University Park, USA-25/05/15,06:43:31) To determine the stress and strain states of the guitar neck, guitar geometry was modeled in CATIA. This was imported in ABAQUS to finite element analysis. In the preprocessing stage, there were introduced the specific parameters of the guitar neck and the guitar body. Table 1 summarizes the elastic characteristics of guitar components taken from literature [6, 7]. In this study, was investigated the statical behaviour of guitar in case of application of concentrated force. The force was applied successively on the frets, simulating down force of strings during playing as can be seen in Fig 2. The intensity of force (F) was varied F=45; 60; 75; 90 N. From the point of view of boundary conditions was considered that the neck was simple supported and end of guitar body was fixed (Fig. 3). Table 2 summarized the values of von Mises stress and maximum displacement for each position of force on the frets obtained with finite element analysis. It was found that in the joint of the neck of the guitar body, applying a force of 45 N produces von Mises stresses greater than those permitted for wood (10\u202612 MPa)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.21-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.21-1.png",
        "caption": "FIGURE 8.21",
        "texts": [
            " ditch in one complete rotation of the cam. Note that since the radius of arc AB is 7.65 in. the cam causes the quarter suspension to travel roughly 41.8 in. (3.48 ft) in one complete rotation. Since the profile cam rotates two complete cycles in one second, the suspension travels about 6.96 ft/sec (i.e., 4.74 MPH), which is very slow. There are nine bodies in thismotionmodel, including the groundbody. There are two rigid (no symbol) joints, three pin joints, eight ball joints, and one cylinder joint, as shown in Figure 8.21. A servomotor that rotates the profile cam for 1 sec was added to conduct a kinematic analysis. Several measures are critical in determining the pros and cons of the suspension design. Among them, the most important one is probably the camber angle. We will show the camber angle results momentarily. Note that the camber angle is defined as the rotation of the upright along the X-axis ofWCS (World Coordinate System, which is the reference frame of the motion model). First, we look at the shock travel distance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000636_chicc.2015.7260198-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000636_chicc.2015.7260198-Figure1-1.png",
        "caption": "Fig. 1: The structure of quad-rotor UAV.",
        "texts": [
            " In this paper, a T-S fuzzy controller is designed for stabilizing the attitude system. The paper is organized as follows: Section 2 presents the complete dynamic system, and derives a simplified dynamic model with some assumptions. Controller for the quad-rotor UAV is designed in Section 3. Simulation results are presented in Section 4. Finally, in Section 5, some conclusions are given. The quad-rotor UAV has four rotors, the front and rear rotors rotate in a counter-clockwise direction while the left and right rotors rotate in a clockwise direction as show in Fig. 1. The quad-rotor system has six freedoms, they concerned with the position and attitude of the quad-rotor UAV. Define the attitude angle vector \u03b7 = [\u03c8, \u03b8, \u03c6]T , where \u03c8 is the yaw angle, \u03b8 is the pitch angle and \u03c6 is the roll angle. The inertia matrix of quad-rotor UAV I = diag(Ix, Iy, Iz), where Ix, Iy, Iz are moments of inertia correspond to the respective axes. And f1, f2, f3, f4 are the thrust produced by corresponding motors. We introduce the complete dynamic model derived with Euler-Lagrange approach [4]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002263_meacs.2014.6986869-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002263_meacs.2014.6986869-Figure3-1.png",
        "caption": "Fig. 3. Bearing Scheme",
        "texts": [
            " 978-1-4799-6221-1/14/$31.00 \u00a92014 IEEE IV. COMPUTING EXPERIMENT A. Description a/t est data Thus, within the available databases twenty one bearing conditions can be categorized: 1) normal; 2-5) defect of rolling body; 6-9) inner track defect; 10-21) outer track defect. Computing experiment was carried out on the basis of database of vibration signals, as described in [5]. This database contains records of vibration signals of rolling bearings, containing defects (on outside and inside track, and on the rolling body (Fig. 3), and without them. The sizes of defects are 0.007, 0.014, 0.021, and 0.028 inches in diameter. The defects of the outer track are stationary, therefore, the position of the defect relatively the load zone of the bearing influences the vibration signal, generated by the bearing. To investigate the influence of this effect, the defects were applied on the outer track in three positions. The scheme of the stand to record vibration signals is shown in Fig. 4. The test bearing is fixed on the motor shaft"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001741_iccas.2015.7364959-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001741_iccas.2015.7364959-Figure1-1.png",
        "caption": "Fig. 1 Geometry of missile and target in I-frame",
        "texts": [],
        "surrounding_texts": [
            "As studies of two-dimensional impact angle control (lAC) guidance laws have been investigated, various works have been done previously [1-4] . These laws studied as the guidance toward stationary targets [1, 2] and moving targets with constant speed [3] , as well as maneuvering targets [4] . The form of these guidance laws consists of the biased PN, which has the two-dimensional proportional navigation (PN) for target interception and the biased term is to control desired impact angle. Additionally, guidance commands of a missile are calculated independently by dividing three-dimensional spaces into two-dimensional planes; pitch and yaw planes. It is possible to calculate the separate guidance commands with pitch and yaw planes due to the simple approximation for the uncomplicated motion of missile. However, it has a limit for calculation of the guidance commands in the case of sophisticated motion. In this paper, we propose the three-dimensional lAC guidance law with the biased PN and navigation constant based on the optimal lAC guidance law in Ref. [4] . The three-dimensional PN is described by the line of sight (LOS) and the missile's velocity from the geometry of missile and targets defined as three dimensional spaces [5] , and the biased term is expressed by using quaternions about attitude of the missile and desired impact angles. 2. THREE-DIMENSNIOAL IMPACT ANGLE CONTROL 2.1 Problem Statement To set the three-dimensional lAC, the dynamics model of Eqs. (1) to (3) is defined in ECI (Earth Centered Inertial) reference frame; I-frame, Fig. I shows the geometry of missile and target in I-frame. The model of missile and target are assumed as point mass, and the guidance law is applied to the set of moving missile and (Ar = 0 because it cannot be calculated from the missile's velocity vector. The subscripts mean that 'M' is missile, 'T' is target, 'TM' is relative state of missile and target, 'f' is final state of missile. ZH = (I) V,, = ax\" ' V,, = a,,, ' V,, = a,,, (2) I/J\" = O. 8\" = tan-' (v:, I \ufffdV,: +17,.: ) . If/\" = tan-' (17\" IV,,) (3) Above dynamics model, the two-dimensional optimal lAC is expressed by Eq. (4) and is shape of the biased PN. 2-D lAC fixed target. The missile's roll angle is assumed as Fig. 2 Approach to formation of three-dimensional lAC 452 978-89-93215-09-0/15/$31.00 @lCROS The PPN (pure PN) and the TPN (true PN) are commonly used in proportional navigation guidance laws. The PPN consists of navigation gain (N), LOS rate (Q), missile's velocity (VIi)' then guidance commands of the PPN are perpendicular to the missile's velocity vector. Otherwise, guidance commands of the TPN using closing velocity (\ufffd.) of the missile and target are perpendicular to the LOS vector. In contrast to the TPN, the PPN's commands are calculated by using LOS rate not to need target' velocity and it is easy to apply by using the measuring device related with LOS angle in effect. Thus, the PPN is considered as base guidance law for the three-dimensional lAC guidance law. In three-dimensional spaces, the PPN can be expressed by vector geometry in Fig. 3. The PPN guidance commands are set of vector product of LOS rate and missile's velocity vector in Eq. (5) and it means that the commands are perpendicular to the missile's velocity. (5) where N is navigation constant, Q = RTM X V 'M is \u2122 IR- .. 12 1:11 LOS rate vector, P\" is missile's velocity vector. 2.3 Euler Axis and Rotation Angle of Quaternion In general, the relative relation between a missile and a target can be presented by coordinate transformation using Euler direction cosine matrix (DCM). In use of Euler DCM, it occurs gimbal lock, which is the problem losing degree of freedom of one or two rotation axes caused by the overlap of two or three axes during the rotation transformation. To avoid gimbal lock, the rotation with quatemion is used. Fig. 4 and Eqs. (6) to (7) show the relation of Euler axis i and rotation angle f.1 about quaternion. The components of i are direction cosines, so i is regarded as normal vector of a plane made up the quaternion. i = cos at + cos fJ] + cos rk (6) q = qo + q/ + q) + q3k f.1 . f.1= cos-+sm-A 2 2 (7) The quaternions can be obtained from the state of missile's attitude and impact angle in reference coordinate, in addition, the quaternion about current state and final state can be calculated by using the relation of two quatemions. Fig. 5 shows the relation of two quaternions qj lf, q : and one quaternion q,;, In Eq. (8), the quaternion q ;f is calculated from the missile's attitude (V.,f, O,f' tf\" ) about 321 rotation [6] . - If/H 0.,1 \u00a2H . If/,H . o,H . <PH q = cos-cos-cos-+sm-sm-sm-, 2 2 2 2 2 2 _ IffH B\" . t/J\" . Iff\" . B\" 1/1\" q, = cos-cos-sm--sm-sm-cos2 2 2 2 2 2 _ If/H \u2022 BH <PH . If/'H BH . <PH q. = cos-sm-cos-+sm-cos-sm2 2 2 2 2 2 _ . IffH B\" t/J\" IffH . B\" . 1/1\" q. = sm-cos-cos--cos-sm-sm-. 2 2 2 2 2 2 (8) Above formulation, quaternion q ; about final state also can be calculated. With two quaternions q;'l and q ;, quaternion q,\ufffd is expressed by Eq. (9). (9) The quaternion q .\ufffd should be transformed by I-frame by the reason why reference frame using in this paper is I-frame. Eqs. (10) to (11) shows the quaternion (q \ufffd) l transformed into I-frame. ( q,\ufffd ) 1 = q: [ q;'l J = q() + PI = qu + q/ + q) + q;k In addition, the rotation angle and Euler axis of are in Eqs. (12) to (13) f.1 = 2 cos -1 q ll li = cos at + cos f3 J + cos rk 1 _ = r:----; PI \" I -q() 2.4 3D Impact Angle Control (10) (11) (12) (13) The three-dimensional lAC guidance commands are expressed by Eq. (14) from the three-dimensional PPN and the quaternion relation as mentioned above. U lAC = N QnJ X \ufffdl + Nf \ufffd (ll X \ufffd1 ) . f.1 t go (14) First one means the three-dimensional PPN, and the other is the biased term for control the desired impact angles. In this lAC, to correspond current state of missile's attitude with final state of desired impact angles, it controls the rotation angle f.1 of quaternion converging to o. It also forms in the velocity vector included in quaternion plane to use vector product of missile's velocity vector and Euler axis of quaternion ."
        ]
    },
    {
        "image_filename": "designv11_64_0000401_978-3-319-00065-7_8-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000401_978-3-319-00065-7_8-Figure2-1.png",
        "caption": "Fig. 2 Phase plane switching control strategy [8]. Brake activation can be fine-tuned using the parameter h. During the active state (b \u223c c, d \u223c e), the brake produces opposing torque as outlined in Equation (2) to compensate overshoots and undershoots. Proper selection of h improves control performance by selectively deactivating the brake when the joint accelerates to the set point, for example in a \u223c b.",
        "texts": [
            " 1), when the brake is used in place of the DC motor, we control the brake using a phase switching controller [8] as follows: \u03c4m = Ked |\u03b8\u0307 |\u03b5 (2) where, Ked is the damping gain for the joint velocity \u03b8\u0307 and \u03b5 = 0 or 1 for the off and on state for the brake, respectively. The phase plane switching surfaces were constructed similarly as [8]. The control strategy allows the user to fine-tune when the brake needs to activate such that fast dynamic response can be achieved despite the brake\u2019s dissipative nature. Fig. 2 illustrates how each switching surface parameter affects the dynamics of the system. The parameters were empirically found as h =\u221250 and Ked = 0.5 with the error defined as e = qd \u2212 q. Note that a linear model relating torque and current is used for the control of brakes. Interestingly, the brake does not transfer torque viscously in shear mode [13]. This means the brake torque is practically determined by a magnetic field, i.e., an input current, rather than an input velocity unless the velocity is significantly low (0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.24-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.24-1.png",
        "caption": "Fig. 4.24 Gravitational field acting on a body i",
        "texts": [
            " In this system, it is possible to identify the main types of forces described above. In a broad sense, the forces in the multibody systems can be divided into two main groups, namely external and internal forces. If the forces are located inside of the boundaries of the system, they are called internal, otherwise the forces are named external forces. Figure 4.23 illustrates a MBS in which the internal and external forces can easily be identified. The definition of the system boundaries depend on the system in analysis. Figure 4.24 illustrates a body i acted upon by a gravitational field in the negative y direction. The choice of the negative y direction as the direction of gravity is totally arbitrary. However, in the present work, the gravitational field will be considered to be acting in this direction unless indicated otherwise. The force due to gravitational field can be written as FG \u00bc mig~uy \u00f04:93\u00de where mi is the mass of body i and g is gravity acceleration. In the present work the acceleration due to gravity assumes a value equal to 9"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002595_j.ifacol.2015.09.678-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002595_j.ifacol.2015.09.678-Figure1-1.png",
        "caption": "Fig. 1. Three in-line array tethered spacecraft formation",
        "texts": [
            " The rest of the paper is organized as follows: Section 2 focuses on the system m del and its simplification. Sections 3 and 4 present for ation control and FTC respe tively. Section 5 gives simulation results, followed by a conclusion in Section 6. 9th IFAC Symposium on Fault Detection, Supervision and Safety of Technical Processes September 2-4, 2015. Arts et M\u00e9tiers ParisTech, Paris, France Copy ight \u00a9 2015 IFAC 1128 Hao Yang et al. / IFAC-PapersOnLine 48-21 (2015) 1128\u20131133 1129 Consider a three-inline array tethered spacecraft formation as shown in Fig.1. The X \u2212 Y coordinate denotes the desired rotating coordinate with respect to the center of mass (c.m.) of center spacecraft 0. For i = 1, 2, the coordinates with subscript i denotes the shifted reference coordinate of X \u2212 Y to measure the angles \u03b8i and \u03c6i. r denotes the distance from the c.m. of spacecraft i to its tether attachment point, Fi and ui are external control on the c.m. of spacecraft i, where Fi is the force due to thruster firing while ui is the torque generated by a reaction wheel assembly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000914_wcica.2014.7053421-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000914_wcica.2014.7053421-Figure1-1.png",
        "caption": "Fig. 1 Geometrical relationship of coordinate systems The model of probe in the process of landing on asteroid is shown in Fig.1, three coordinate systems displayed in the",
        "texts": [
            "00 \u00a92014 IEEE Proceeding of the 11th World Congress on Intelligent Control and Automation Shenyang, China, June 29 - July 4 2014 4213 the x,y,z axis compose the right-handed coordinate system. Let the landing site system fixed on the asteroid with the origin coinciding with the landing site ( ), and the x-axis coinciding with the unit vector from the landing site to the point , and the z-axis is the normal line of the plane that is determined by the three points ( ), the x,y,z axis composed the righthanded coordinate system as shown in Fig.1. If is the unit directional vectors of the landing site coordinate system, the landing site coordinate system can be constructed as followed: (1) \u00d7 \u00d7 (2) (3) Where , ( ) denotes the vector from the center of the asteroid to the point ( ), the superscript \u2018a\u2019 is the fixed-body coordinate system . So, the transformation matrix from the landing site system to the fixed-body coordinate system can be written as followed: (4) The camera coordinate system is fixed on the camera, and the plane of image is defined as the plane of o-xy, z axis is coordinate with the optical axis of camera"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure11-1.png",
        "caption": "Figure 11. Torque limiter (TGB20-HC, Tsubakimoto Chain Co.) [14].",
        "texts": [
            " After Plate A is locked, each Claw B slides along each Guide Hole A of Plate A by the rotation of Plate B (Fig. 10(b)) and then one of three Claws B is hooked to the inner teeth and rotates Plate C (Fig. 10(c)). By the rotation of Plate C, Pin C switches off and each Claw C moves along each Guide Hole C (Fig. 10(d)). After that, four Claws C simultaneously mesh with Ratchet Wheel C and thus Shaft C is locked. The lock of Shaft C is released by rotating Shaft C in a direction opposite to the direction locked by the safety device. The torque-based safety device consists of a commercial torque limiter (Fig.11, TGB20-HC, Tsubakimoto Chain Co., see [14] for more details) and a switch (Switch Z). As shown in Fig. 12, Switch Z is installed at the position of being pressed by Plate Z of the torque limiter when Plate Z slides along Hub of the torque limiter. Switch Z can interrupt electric power supply to all motors of the robot. Shaft A is connected to Hub. Hub has a threaded portion and Adjusting Nut is attached to Hub. Balls are respectively inserted into Holes of Center Flange and set in V-shaped Pockets of Hub"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002232_icma.2014.6885882-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002232_icma.2014.6885882-Figure2-1.png",
        "caption": "Fig. 2. The CAD model of general robotic arm.",
        "texts": [
            " In this paper, we try to propose a new kind of preoperative positioning method for a combined surgical robot which is consisted of three same robotic arms, the mid arm is installed with a laparoscope and the other two sides are installed with micro instrument devices; the two sets of solutions of the instrument arms\u2019 preoperative positioning angles are calculated by using numerical computation method, we use the projection length & area determination conditions and the percentage of collaboration workspace to achieve the best preoperative positioning angles of the two instrument arms. This paper focuses on the instrument arms\u2019 preoperative positioning problem of the MISR as shown in Fig. 1, the slave arms of MISR are made up of two instrument arms and one laparoscope arm. 1269978-1-4799-3979-4/14/$31.00 \u00a92014 IEEE The three robotic arms are the same 8 DoFs general robotic arm except for the devices on the end, as shown in Fig. 2, the first four joints are passive and the others are active. Before the operation, the surgical assistants will place the three robotic arms to the appropriate position and orientation, therefore the trocars\u2019 position are fixed to the desired position that is determined by the actual operation requirements. As a multi-arms robot system, the three arms can not be intervention with each other; the two instrument arms have to be placed into the proper positions so that they can not be intervention between themselves and their collaboration workspace can be enough for operation, the interspace between the two instrument arms must be suited for laparoscope arm\u2019s preoperative position"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.10-1.png",
        "caption": "FIGURE 6.10",
        "texts": [
            " Also, although in the sketch the swing arms are shown as a \u2018plausible\u2019 mechanical arrangement (i.e. not overlapping) this also need no be so; in general, contact between elements is not modelled for vehicle dynamics studies and in general the instant centres are widely spaced and not necessarily within the physical confines of the vehicle body. The swing armmodel has the advantage over the equivalent roll stiffness model in that the heave and pitch ride behaviour can be included. The model based on linkages as shown in Figure 6.10 is the model that most closely represents the actual vehicle. This sort of vehicle model is the most common approach adopted by MBS software users in the automotive industry, often extending the model definition to include full nonlinear bush characteristics. A simplification of a model based on linkages is to treat the joints as rigid and generate a kinematic representation of the suspension system. As described in Chapter 4 a double wishbone arrangement is typical of a suspension system that can be modelled in this way and used for handling simulations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000726_1.4031738-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000726_1.4031738-Figure3-1.png",
        "caption": "Fig. 3 MR frames (top view)",
        "texts": [],
        "surrounding_texts": [
            "VHB I;Pdm\n\u00bc uI;Pdm\nvI;Pdm wI;Pdm\n0 @\n1 A HB\n\u00bc u v w\n0 @ 1 A b \u00fe d\ndt\nxdm ydm zdm\n0 @\n1 A HB\n\u00fe p q r\n0 @ 1 A b\nxH yH zH\n0 @ 1 A b \u00fe xdm\nydm zdm\n0 @\n1 A HB 0 B@ 1 CA (18)\nwhere we have used\nT HB\u00f0 \u00deo\nVN VE VZ\n0 @ 1 A o\n\u00bc T HB\u00f0 \u00deb:Tbo: VN VE\nVZ\n0 @ 1 A o\n(19)\ntogether with T HB\u00f0 \u00deb \u00bc I, and Tbo: VN VE\nVZ\n0 @ 1 A o\n\u00bc u v w\n0 @ 1 A b\nfrom the Nomenclature. Now substituting Eq. (9) into Eq. (18), we can obtain an expanded expression for VHB\nI;Pdm\nuHB I;Pdm \u00bc u\u00fe XMR sin wbl eL \u00fe eP \u00fe cos fbl eF \u00fe rdm cos bbl\u00f0 \u00de\u00bd cos wbl cos hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de \u00fe rdm sin bbl sin hbl\u00bd \u00f0 \u00de \u00fe _fbl eF \u00fe rdm cos bbl\u00f0 \u00de cos wbl sin fbl sin wbl cos hbl cos fbl\u00bd \u00fe _bblrdm cos wbl cos fbl sin bbl \u00fe sin wbl cos hbl sin fbl sin bbl cos bbl sin hbl\u00f0 \u00de\u00bd \u00fe _hbl sin wbl sin hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de rdm sin bbl cos hbl\u00bd \u00fe q zH rdm cos hbl sin bbl \u00fe eF \u00fe rdm cos bbl\u00f0 \u00desin fbl sin hbl\u00f0 \u00de r yH C cos wbl cos hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de \u00fe rdm sin bbl sin hbl\u00f0 \u00de \u00fe C sin wbl eL \u00fe eP \u00fe cos fbl eF \u00fe rdm cos bbl\u00f0 \u00de\u00f0 \u00de\u00f0 \u00de\n(20)\nvHB I;Pdm \u00bc v\u00fe XMRC eL \u00fe eP\u00f0 \u00decos wbl \u00fe rdm sin wbl sin bbl sin hbl \u00fe eF \u00fe rdm cos bbl\u00f0 \u00de cos wbl cos fbl \u00fe sin wbl cos hbl sin fbl\u00f0 \u00de\n_fblC eF \u00fe rdm cos bbl\u00f0 \u00de cos wbl cos fbl cos hbl \u00fe sin wbl sin fbl\u00bd \u00fe _bblrdmC cos wbl cos hbl sin fbl sin bbl cos wbl cos bbl sin hbl sin wbl cos fbl sin bbl\u00f0 \u00de \u00fe _hblC cos wbl sin hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de rdm sin bbl cos hbl\u00bd p zH rdm cos hbl sin bbl eF \u00fe rdm cos bbl\u00f0 \u00desin fbl sin hbl\u00f0 \u00de\u00f0 \u00de \u00fe r xH cos wbl eL \u00fe eP \u00fe cos fbl eF \u00fe rdm cos bbl\u00f0 \u00de\u00f0 \u00de \u00fe sin wbl cos hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de \u00fe rdm sin bbl sin hbl\u00f0 \u00de\u00f0 \u00de\u00f0 \u00de\n(21)\nwHB I;Pdm \u00bc w\u00fe _fbl cos fbl sin hbl eF \u00fe rdm cos bbl\u00f0 \u00de _bblrdm cos bbl cos hbl \u00fe sin bbl sin fbl sin hbl\u00f0 \u00de \u00fe _hbl rdm sin hbl sin bbl \u00fe eF \u00fe rdm cos bbl\u00f0 \u00desin fbl cos hbl\u00bd \u00fe p yH C cos wbl cos hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de \u00fe rdm sin bbl sin hbl\u00f0 \u00de \u00fe C sin wbl eL \u00fe eP \u00fe cos fbl eF \u00fe rdm cos bbl\u00f0 \u00de\u00f0 \u00de\u00f0 \u00de q xH cos wbl eL \u00fe eP \u00fe cos fbl eF \u00fe rdm cos bbl\u00f0 \u00de\u00f0 \u00de sin wbl cos hbl sin fbl eF \u00fe rdm cos bbl\u00f0 \u00de \u00fe rdm sin bbl sin hbl\u00f0 \u00de\u00f0 \u00de (22)\nwith the total blade pitch angle given by [59]\nhbl \u00bc h0 \u00fe h1c cos wbl \u00fe wPA\u00f0 \u00de \u00fe h1s sin wbl \u00fe wPA\u00f0 \u00de \u00fe ht;rdm\nK hblbbl\u00f0 \u00debbl K hblfbl\u00f0 \u00defbl (23)\nand the blade pitch component due to blade twist given by\nht;rdm \u00bc rdm\nhwash\nRbl\n(24)\nNote also, as stated in Sec. 3.1, we neglect any effects due to rapid pitch changes, e.g., dynamic stall effects. Hence, in the sequel we will assume that _hbl _bbl;\n_hbl _fbl , and _hbl XMR. Consequently, in the sequel we will also assume to have _hbl \u2019 0 in Eqs. (20)\u2013(22).\n3.4 Flap\u2013Lag Equations of Motion. Since the early 1950 s it is known that including flapping dynamics in a helicopter flight model could produce limitations in rate and attitude feedback gains [60]. Further, for helicopter directional axis control, blade lead\u2013lag dynamics ought to be considered for control system\ndesign [61]. Indeed, it is well known that blade lead\u2013lag produces increased phase lag at high frequency, in the same frequency range where flapping effects occur [62], and that control rate gains are primarily limited by lead\u2013lag\u2013body coupling [62,63]. Now, in terms of blade flap\u2013lag modeling, a foundational contribution was given in Ref. [59], where derivations of the coupled flap\u2013lag equations of motion for a rigid articulated rotor, for the (F\u2013L\u2013P), (F\u2013P\u2013L), and (L\u2013F\u2013P) sequences, were laid out. The purpose of our work is to present a model for a new hinge arrangement, i.e., the (P\u2013L\u2013F) sequence, which is much more useful for modeling the rotor dynamics of a small-scale helicopter. The equations presented in the sequel (obtained by the Lagrangian method [64]) are valid for a single articulated rotor with hinge springs and viscous dampers. Compared to Ref. [59], our approach retains all three hinges physically separated and works also for both clockwise (CW) and counterclockwise (CCW) rotating MRs. Further, full coupling between vehicle and blade dynamics is modeled. Now from Lagrangian theory, we have\nd\ndt\n@KE @ _fbl\n! @KE\n@fbl\n\u00bc Qfbl (25a)\nJournal of Dynamic Systems, Measurement, and Control JANUARY 2016, Vol. 138 / 011010-5\nDownloaded From: http://dynamicsystems.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use",
            "d dt @KE\n@ _bbl\n! @KE\n@bbl\n\u00bc Qbbl (25b)\nwith KE the kinetic energy of a blade, bbl and fbl the blade flap and lag angles, and Qbbl and Qfbl the generalized forces. The latter include the effect of gravity, aerodynamics, and spring damping and stiffness and are given by\nQfbl \u00bc Qfbl;G \u00fe Qfbl;A \u00fe Qfbl;D \u00fe Qfbl;S (26a)\nQbbl \u00bc Qbbl;G \u00fe Qbbl;A \u00fe Qbbl;D \u00fe Qbbl;S (26b)\nThe kinetic energy of a single rotor blade is given by\nKE \u00bc 1\n2\n\u00f0Rbl\n0\nVHB I;Pdm >:VHB I;Pdm dm (27)\nwith VHB I;Pdm\ncomputed in Eq. (18), and the limits of integration are from the flap hinge, to the blade tip. The kinetic energy inboard of the flap hinge is neglected in our model since assumed small in the case of small-scale UAVs. Next, we provide the procedure for\nthe blade lead\u2013lag equations (Eq. (25a)), the blade flap equations (Eq. (25b)) follow a similar reasoning and are thus omitted. Now, we rewrite the first term on the left-hand side (LHS) of Eq. (25a) as\nd\ndt\n@KE @ _fbl\n! \u00bc d\ndt\n@\n@ _fbl\n1\n2\n\u00f0Rbl\n0\nVHB I;Pdm >:VHB I;Pdm dm\n! (28)\nAnd since the limits of integration are constant, Eq. (28) is equivalent to (using Leibniz\u2019s integral rule)\n1\n2\n\u00f0Rbl\n0\nd\ndt\n@\n@ _fbl\nVHB I;Pdm >:VHB I;Pdm dm (29)\nNext using the chain rule, Eq. (29) is equivalent to\n1\n2\n\u00f0Rbl\n0\nd dt 2 VHB I;Pdm >: @\n@ _fbl\nVHB I;Pdm\n! dm\n\u00bc \u00f0Rbl\n0\nVHB I;Pdm\n>: d\ndt\n@\n@ _fbl\nVI I;Pdm\n!HB 2 4\n\u00fe d\ndt VI I;Pdm\n> HB\n: @\n@ _fbl\nVHB I;Pdm\n# dm (30)\nwith again the following convention for the time-derivatives:\nd=dt\u00f0 \u00de @=@ _fbl VI\nI;Pdm\nHB\nsignifies the time-derivative, with\nrespect to inertial frame FI, of vector @=@ _fbl VI;Pdm\n, subsequently projected onto frame FHB. Using Eq. (16), these derivatives can also be expanded as follows:\nd\ndt\n@\n@ _fbl\nVI I;Pdm\n!HB\n\u00bc d\ndt\n@\n@ _fbl\nVHB I;Pdm\n!HB\n\u00fe p q r\n0 @ 1 A b\n@\n@ _fbl\nVHB I;Pdm\n(31)\nd dt VI I;Pdm\n> HB\n\u00bc d\ndt VHB I;Pdm\n> HB \u00fe p q r\n0 @ 1 A b\nVT I;Pdm\nHB\n(32)\nFig. 4 MR frames (side view)\n011010-6 / Vol. 138, JANUARY 2016 Transactions of the ASME\nDownloaded From: http://dynamicsystems.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use",
            "Next, for the second term on the LHS of Eq. (25a) we get\n@KE @fbl \u00bc @ @fbl 1 2\n\u00f0Rbl\n0\nVHB I;Pdm >:VHB I;Pdm dm (33)\nAgain since the limits of integration are constant, and using the chain rule, Eq. (33) reduces to\n@KE @fbl\n\u00bc \u00f0Rbl\n0\nVHB I;Pdm\n>: @\n@fbl\nVHB I;Pdm dm (34)\nNow, through the use of a symbolic math toolbox, an analytic expression for the LHS of Eq. (25a) may readily be obtained, i.e., by utilizing the expression obtained for VHB I;Pdm in Eq. (18) and inserting it, together with the derivatives d=dt\u00f0 \u00deVHB I;Pdm ; @=@\u00f0 \u00defblV HB I;Pdm ; and @=@ _fbl VHB I;Pdm into Eqs. (30)\u2013(32) and (34). The blade flap equation Eq. (25b) follows a similar procedure and will also require the computation of @=@bbl\u00f0 \u00deVHB I;Pdm and\n@=@ _bbl VHB I;Pdm . Finally, using a symbolic math toolbox, the combined equations (Eqs. (25a) and (25b)) may be rearranged as the following four-state nonlinear flap\u2013lag equations of motion:\nd\ndt\n_bbl\n_fbl\nbbl\nfbl\n0 BBBBBBB@ 1 CCCCCCCA \u00bc A 1: B: _bbl _fbl bbl\nfbl\n0 BBBBBBB@ 1 CCCCCCCA \u00fe Qbbl F1 Qfbl F2 0\n0\n0 BBBBBB@\n1 CCCCCCA 0 BBBBBBB@ 1 CCCCCCCA (35)\nwith the following A and B matrices:\nA \u00bc\nIb 0 0 0 0 e2 F:Mbl \u00fe 2eF:C0 \u00fe Ib 0 0 0 0 1 0\n0 0 0 1\n2 664\n3 775\nB \u00bc\n0 B12 0 0 B21 0 0 0\n1 0 0 0\n0 1 0 0\n2 664\n3 775\n(36)\nwith Mbl, C0, and Ib defined as (refer also to the Nomenclature)\nMbl \u00bc \u00f0Rbl\n0\ndm C0 \u00bc \u00f0Rbl\n0\nrdm:dm \u00bc Mbl:yGbl\nIb \u00bc \u00f0Rbl\n0\nr2 dm:dm \u00bc Mbl:\nR2 bl\n3 (37)\nWe stress here that Eq. (35) is a nonlinear representation since the scalars B12 and B21 in Eq. (36), and F1 and F2 in Eq. (35) are (nonlinear) functions of _fbl; bbl; fbl . Space restrictions preclude a reprint of the lengthy expressions B12, B21, F1, and F2, these can be consulted in Appendix E of Ref. [65].\n3.4.1 Flap Angle as a Fourier Series. Blade motion is 2p periodic around the azimuth and may hence be expanded as an infinite Fourier series [54,66]. Now for full-scale helicopters, it is well known that the magnitude of the flap second harmonic is less than 10% the magnitude of the flap first harmonic [33,54]. We assume that this is also the case for small-scale helicopters and hence we neglect second and higher harmonics in the Fourier series. This gives\nbbl wbl\u00f0 \u00de \u2019 b0 \u00fe b1c cos wbl \u00fe b1s sin wbl (38)\nwith wbl the blade azimuth angle. This harmonic representation of the blade motion defines the rotor tip-path-plane (TPP), resulting in a so-called cone-shaped rotor. The nonperiodic term b0 describes the coning angle, and the coefficients of the first harmonic b1c and b1s describe the tilting of the rotor TPP, in the longitudinal and lateral directions, respectively. All three angles may readily be obtained through standard least-squares [67]. Now in steady-state rotor operation, the flap coefficients b0, b1c; and b1s may be considered constant over a 2p blade revolution. Obviously, this solution would not be adequate for transient situations, such as maneuvering [68], hence in our model we compute, for each new blade azimuth, the instantaneous TPP angles. With regard to TPP dynamics, three natural modes can be identified, i.e., the so-called coning, advancing, and regressing modes. In general, the regressing flapping mode is the most relevant when focusing on helicopter flight dynamics, as it is the lowest frequency mode of the three, and it has a tendency to couple into the fuselage modes [53,62,69].\n3.4.2 Virtual Work and Virtual Displacements. The determination of the generalized forces Qfbl and Qbbl in Eqs. (26a) and (26b) requires the calculation of the virtual work of each individual external force, associated with each respective virtual flapping and lead\u2013lag displacements [59]. Let FXi ;FYi ; and FZi\nbe the components of the ith external force Fi, acting on BE dm in frame FHB, then the resulting elemental virtual work done by this force, due to the virtual flapping and lag displacements @bbl and @fbl, is given by\ndWi \u00bc FXi dxdm \u00fe FYi dydm \u00fe FZi dzdm (39)\nwith\ndxdm \u00bc @xdm\n@bbl\n@bbl \u00fe @xdm\n@fbl\n@fbl (40a)\ndydm \u00bc @ydm\n@bbl\n@bbl \u00fe @ydm\n@fbl\n@fbl (40b)\ndzdm \u00bc @zdm\n@bbl\n@bbl \u00fe @zdm\n@fbl\n@fbl (40c)\nNow summing up the elemental virtual work, over the appropriate blade span, results in the total virtual work Wi, due to external force Fi, as\nWi \u00bc \u00f0Rbl\n0\nFXi\n@xdm @bbl \u00fe FYi @ydm @bbl \u00fe FZi @zdm @bbl @bbl\n\u00fe \u00f0Rbl\n0\nFXi\n@xdm @fbl \u00fe FYi @ydm @fbl \u00fe FZi @zdm @fbl @fbl (41)\nWhich is set equivalent to\nWi \u00bc Qbbl;i:@bbl \u00fe Qfbl;i:@fbl (42)\nThe virtual displacement, in frame FHB, of a BE dm, located at a distance rdm outboard of the flap hinge, is obtained using Eqs. (40) and (9) as follows:\ndxdm dydm dzdm\n0 @\n1 A HB\n\u00bc rdm:dPHB b;r :@bbl \u00fe dPHB f; r \u00fe rdm:dPHB f;r h i :@fbl (43)\nJournal of Dynamic Systems, Measurement, and Control JANUARY 2016, Vol. 138 / 011010-7\nDownloaded From: http://dynamicsystems.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use"
        ]
    },
    {
        "image_filename": "designv11_64_0000493_iecon.2014.7049076-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000493_iecon.2014.7049076-Figure5-1.png",
        "caption": "Fig. 5. Prototype PMSMs used in this paper",
        "texts": [
            " (34) and Eq. (35), respectively. nK nL e qx (33) ni nin T nini q e s T 1A (34) niRnv sy (35) Similar to the case of standstill, Eq. (26) Eq. (32) are the recursive formulas for solving Lq and Ke. IV. EXPERIMENTAL SETUPS AND RESULTS The PMSM drive algorithms developed in this paper are implemented using a low cost 16-bit fixed-point microcontroller dsPIC33FJ32MC304 manufactured by Microchip Technology. Fig. 4 shows the prototype of the sensorless FOC drive board of the PMSM developed in this paper. Fig. 5 shows the photographs of the surface-type and interior-type PMSMs used in the experiment. Note that the values of the motor parameters, rated power, and maximum operation speeds of these two motors are different. Several experiments have been conducted to verify the effectiveness of the proposed sensorless FOC drive with automatic PMSM parameter identification of the surface-type and interior type PMSMs. Fig. 6 shows the measured current waveforms for the surface-type and interior type PMSMs using the proposed startup procedure with automatic PMSM parameter identification"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000734_978-3-319-08422-0_4-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000734_978-3-319-08422-0_4-Figure4-1.png",
        "caption": "Fig. 4 Inverted Pendulum Controlled by DC Motor",
        "texts": [
            " u\u00f0bx, t\u00de \u00bc R 1\u00f0bx\u00deB0 \u00f0bx\u00de\u00bdPc\u00f0bx, t\u00debx\u00f0t\u00de g\u00f0bx, t\u00de , \u00f027\u00de where, Pc(x) is a positive-definite solution of the continuoustime State Dependent Riccati Equation Pc\u00f0bx\u00deA\u00f0bx\u00de \u00fe A 0 \u00f0bx\u00dePc\u00f0bx\u00de P\u00f0bx\u00decB\u00f0bx\u00deR 1\u00f0bx\u00deB0 \u00f0bx\u00dePc\u00f0bx\u00de \u00fe C 0 \u00f0bx\u00deQ\u00f0bx\u00deC\u00f0bx\u00de\u00bc 0, \u00f028\u00de and g(x) is a solution of the continuous-time State Dependent non-homogeneous equation g\u00f0bx\u00de\u00bc \u00bdA\u00f0bx\u00de B\u00f0bx\u00deR 1\u00f0bx\u00deB0 \u00f0bx\u00dePc\u00f0bx\u00de 0 1 C 0 \u00f0bx\u00deQ\u00f0bx\u00dez\u00f0bx\u00de, \u00f029\u00de For numerical simulation and analysis, the developed estimation and optimal tracking technique is implemented for noise cancellation for inverted pendulum controlled by DC motor, as shown in Fig. 4. The dynamic equations for system under concern are: V\u00f0t\u00de \u00bc L di\u00f0t\u00de dt \u00fe Ri\u00f0t\u00de \u00fe kb d\u03b8\u00f0t\u00de dt , \u00f030\u00de ml2 d2\u03b8\u00f0t\u00de dt2 \u00bc mglsin\u00f0\u03b8\u00f0t\u00de\u00de kmi\u00f0t\u00de, \u00f031\u00de where, V is the control voltage, L is the motor inductance,i is the current through the motor winding, R the motor winding resistance,kb the motor\u2019s back electro magnetic force constant, \u03b8 the angle of pendulum,m the mass of pendulum, l the length of rod, g the gravitational constant, and km the damping (friction) constant. The system nonlinear state equations can be written in the state dependent form: _x1 _x2 _x3 264 375 \u00bc 0 1 0 g l sin\u00f0x1\u00de x1 0 km ml2 0 kb L R L 266666664 377777775 x1 x2 x3 264 375\u00fe 0 0 1 L 26664 37775u: \u00f032\u00de where: \u03b8 \u00bc x1, _\u03b8 \u00bc x2, i \u00bc x3, V \u00bc u"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002694_9781782421955.869-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002694_9781782421955.869-Figure7-1.png",
        "caption": "Figure 7: Radar charts for efficiency (left) and transmission error (right)",
        "texts": [
            " This function varies modifications within predefined limits, ascertains all possible combinations of up to three sets of modifications and carries out a contact analysis for each variant. At the same time it is possible to vary load as contact behavior strongly depends on load. Figure 6: Sliding speed, local power loss and normal force The results are visualized in clear form in radar charts, allowing several parameters to be reviewed simultaneously. This enables the engineer to determine the optimum combination of modifications for the case at hand. Figure 7 shows two radar charts, the left-hand one depicting efficiency, the right-hand one the transmission error. Here, variant 3:2 would constitute a good compromise exhibiting high efficiency (99.6% instead of 99.06%) and moderate transmission error. 4 SUMMARY The applications shown above represent the various but also different demands in efficiency and thermal rating calculation for gearboxes. Whereas the industrial gearboxes include mainly machine elements for which the calculation methods already established and focus on thermal rating finally, the vehicle industry needs to meet the demands on CO2 reduction targets and hence the efficiency calculation is important"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000981_amc.2014.6823304-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000981_amc.2014.6823304-Figure3-1.png",
        "caption": "Fig. 3. Structure of one leg",
        "texts": [
            " However, vibration of the steering joints occurs if the wheel speed is very low. This is a singularity problem which has still remained open. In this paper, we show three-dimensional forward kinematics model that enables the proper motion generation. In addition, we describe the measures of problem that occurs in the case of the singular configuration avoidance in acceleration level of our robot. II. RobotModeling The robot model and coordinate systems are depicted in Fig.2. And structure of one leg is shown in Fig.3. The link parameters of the robot are listed in Table.I. Each coordinate frame is defined as follows. \u2022 \u03a3W : Coordinates of the world \u2022 \u03a3B : Coordinates of the base link \u2022 \u03a3ci : Coordinates of the contact point of wheel Then, \u2022 W pB, B pci \u2208 R3\u00d71 are position vector of base link with respect to \u03a3W and contact point of each wheel with respect to \u03a3B where i = 1, \u00b7 \u00b7 \u00b7 , k represents the index of the leg. k is the total number of legs (k = 4 for our robot). \u2022 \u03a6 = [\u03c6 \u03b8 \u03c8]T are euler attitude angle of base link"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001149_2014-01-1797-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001149_2014-01-1797-Figure3-1.png",
        "caption": "Figure 3. Overview of hybrid transmission",
        "texts": [
            " 1555 The architecture of the second-generation hybrid system is basically the same as that of the first-generation system, as shown in Figure 1 and Figure 2. The 1M2CL parallel full hybrid system is composed of gasoline engine, one electric motor/ generator, inverter, lithium-ion battery and two mechanical clutches. The motor can be used for several purposes-motoronly driving, engine assist, engine start, regeneration during coasting, and engine-based generation-owing to both the high-output, fast charging/discharging lithium-ion battery and the precise, high-speed motor control. As shown in Figure 3, the torque converter was removed from the transmission and a dry single-plate clutch (clutch-1) with a hydraulic actuator was positioned in its place to achieve both excellent fuel efficiency and a direct acceleration feeling. Owing to complete disconnection of the engine and the motor by clutch-1, the vehicle can be propelled by the motor alone and the region of EV driving has been expanded to even the high-speed range by the high-capacity battery. A wet multiplate clutch (clutch-2) inside the 7-speed automatic transmission is used as a launching element instead of the torque converter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003439_robio.2016.7866382-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003439_robio.2016.7866382-Figure12-1.png",
        "caption": "Fig. 12. The frame configurations in the origami-folding.",
        "texts": [
            " The result of the piecewise polynomials pp = {ppi } is shown in the right column of Fig. 11, where the red lines represent the piecewise polynomials in the x, y, z-axes while the blue lines are the original trajectories. It is found that the piecewise polynomials are modeled appropriately because any noise component are rejected and the essential behaviors are preserved. C. Inverse Kinematics via Mathematical Programming The situation of the origami-folding task for the robotic manipulation system is illustrated in Fig. 12. The resultant piecewise polynomial p\u22c6 i (i = 1, 2, \u22c6 \u2208 {r, l}) is transformed for the trajectory which is appropriately scaled based on two types of origami-paper: P p\u0304F\u22c6 i (t) = lH lF p\u22c6 i (t), (13) where lH = 150, lR = 230 [mm] are the sizes of the origami-paper for a human and the robot finger respectively. The scaled trajectory P p\u0304F\u22c6 i (t) is transformed from the paper frame \u03a3P to the base frame \u03a3B : Bp\u0304F\u22c6 i (t) = BRP P p\u0304F\u22c6 i (t) + BpP , (14) where BpP = [500, 0, \u2212350]T [mm], BRP = I3 represent the position and orientation of \u03a3P with respect to \u03a3P "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003867_978-3-319-03500-0_11-Figure11.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003867_978-3-319-03500-0_11-Figure11.1-1.png",
        "caption": "Fig. 11.1 Wheeled mobile robot",
        "texts": [
            " In the literature, there are few works which makes use of the multiple models based approach for the control of wheeled mobile robots. De La Cruz and Carelli [22] have proposed a switching control for novel tracking and positioning adaptive control of wheeled mobile robots that uses multiple parameter updating laws with different gains. A method that utilizes multiple models of the robot for its identification in an adaptive and learning control framework has been presented by [23]. The model of a wheeled mobile robot (Fig. 11.1) which is subjected to m constraints may be derived as M(q)q\u0308 + C(q, q\u0307)q\u0307 = B(q)\u03c4 + AT (q)\u03bb (11.1) where q \u2208 Rn is generalized coordinates, \u03c4 \u2208 Rr is the input torque vector, \u03bb \u2208 Rm is the vector of constraint forces, M(q) \u2208 Rn\u00d7n is a symmetric positive-definite inertia matrix, C(q, q\u0307) \u2208 Rn\u00d7n is coriolis matrix, B(q) \u2208 Rn\u00d7r is the input transformation matrix, and A(q) \u2208 Rm\u00d7n is the matrix associated with the constraints (Table 11.1). Table 11.1 Model parameters of nonholonomic wheeled mobile robot Parameter Description r Driving wheel radius 2b Distance between two wheels d Distance point Pc from point P0 a Distance from P0 to Pa mc The mass of the platform without the driving wheels and the rotors of the DC motors mw The mass of each driving wheel plus the rotor of its motor IC The moment of inertia of the platform without the driving wheels and the rotors of the motors about a vertical axis through Pc Iw The moment of inertia of each wheel and the motor rotor about the wheel axis Im The moment of inertia of each wheel and the motor rotor about a wheel diameter Assuming that the velocity of P0 is in the direction of x-axis of the local frame and there is no side slip, and considering q = [ x0 y0 \u03d5 ]T , the following constraint with respect to P0 is obtained x\u03070 sin \u03d5 \u2212 y\u03070 cos\u03d5 = 0 (11"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001886_12.2202394-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001886_12.2202394-Figure1-1.png",
        "caption": "Figure 1. Microfluidic technology for toxin avoidance behavior studies on naupli of marine shrimp Artemia franciscana. (A) 3D CAD model of the microfluidic circuitry. The design enables \u201ccaging\u201d of Artemia within a test chamber. (B) Computational fluid dynamics (CFD) simulation of laminar flow inside the designed microfluidic device. (C) Principles of device operation.",
        "texts": [
            " For the acute toxicity tests, concentration-response curves of immobilization were modeled using the Hill model in the Prism 6 and ToxRat Pro. For behavioral analysis, a standard ANOVA model was applied with a Student\u2019s t-test to perform independent comparisons of each toxicant concentration. Every concentration for each hour was compared to the independently run control tests. The Lab-on-a-Chip prototype was a 3D multilayer device that contained a single laser-cut caging chamber for holding of freely swimming crustacean larvae under continuous microperfusion (Figure 1). The chamber was 10 mm in diameter and 1.5 mm in height with two separate inlets and one common outlet channels. Inlets and outlet were interconnected with the main chamber through ten laser-ablated microchannels with a height of proximately 0.2 mm. The engraved microchannels were positioned at the perimeter of the caging chamber (Figure 1). Their dimensions prevented freeswimming specimens to escape from the test chamber. In order to improve mass transfer, inlet and outlet microchannels were ablated in opposite layers to provide a fluid flow across vertical plane of the microchamber. The design also included a separate specimen-loading channel that enabled rapid and efficient loading of test specimens. Sealing of the auxiliary manifold after loading of free-swimming nauplii was realized by injecting a small air bubble (Figure 1). Effective circulation of medium and creation of split fluids zones for behavioral toxicant avoidance tests was a priority for this prototype design. CFD simulations were thus performed to predict flow velocity, pressure drop and uniformity of mass transfer inside the device. Guided by CFD simulations and it was estimated that a flow rate of 2.5 mL/h at each inlet port would result in an average flow velocity of 0.41E-03 m/s to 0.43E-03 m/s, at the input and output, respectively. At those velocities CFD modeling indicated generation of two stable and discrete domains to perform toxin avoidance tests (Figure 1). Proc. of SPIE Vol. 9668 96680Y-3 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 02/04/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx The CFD models were then validated experimentally. Perfusion with a Trypan blue dye confirmed generation of two stable and discrete laminar domains at flow rates of 2.5 mL/h (Figure 2). Repetitive passages of free-swimming Artemia sp. nauplii did not result in destabilization of the fluidic boundary zone or in excessive mixing between discrete domains (Figure 2)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002735_b978-0-12-803081-3.00009-7-Figure9.7-1.png",
        "caption": "FIGURE 9.7",
        "texts": [
            " It is seen that the DQM with small or large initial guesses yields the same results for the problem considered. The reason is that only helical buckling will occur for this problem [4] and the half wave number of the postbuckling mode shape does not change with the increase in applied load; therefore, whether the initial guess \u03b8{ } 0 is large or small does not affect the final results. The DQ results are almost exactly the same as the DSC results presented in Ref. [3]; the formulations validate each other. Figure 9.7 shows the postbuckling configuration at P = 2240 N. The maximum rotation is about 19.133 radians, i.e., umax = 6.09\u03c0. The two plots represent the same postbuckling configuration but viewing at different directions. Note that different units are used in Fig. 9.7 for clarity. The unit in the x and y direction is centimeter and the unit in the z direction is meter. To verify whether the results are valid data or not, \u03b2 \u03b8( )= rd ds/ and the contact force per unit length (Wn) are also computed and plotted. Figure 9.8 shows the variations of b at various applied axial loads. From Fig. 9.8, it is seen that b is small and the absolute values are less than 0.04; therefore, the assumption of \u03b2 \u03b2\u2248sin( ) is held. It is also observed that the pitch of the helix is almost constant for most portions of the cylinder but varies greatly in the middle portion of the cylinder"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.26-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.26-1.png",
        "caption": "FIGURE 8.26",
        "texts": [
            "24, a 150 lb external force pointing upward was applied on the road profile cam to mimic the wheel load due to racecar weight (445 lb) and driver weight (155 lb). An equilibrium analysis was first carried out. The equilibrium state of the racecar was assumed as the initial condition for the dynamic simulation, in which the racecar started in equilibrium on the flat road and then reached the first hump. A spring and a damper were also defined in the dynamic analysis, as shown in Figure 8.25. The physical position of the spring is shown in Figure 8.26. The spring rate and the damping coefficient were 100 lb/in. and 10 lb/(in./sec), respectively. The free length of the spring was 5.5 in. Note that when the shock was fully extended to its maximum length, the spring length was 4 in., which implies a 150 lb preload. The dynamic simulation (Case A) was carried out assuming a racecar speed of 4.74 mph. The shock travel is shown in Figure 8.27. Note that the shock length was allowed to vary between 7.3 in. and 9 in., as shown in Figure 8.28. This means the shock travel obtained in this dynamic simulation (Case A) is acceptable and the design is safe"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000803_ccece.2014.6901107-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000803_ccece.2014.6901107-Figure1-1.png",
        "caption": "Fig. 1. Rotor of an IPMHS Motor",
        "texts": [
            " INTRODUCTION An interior permanent magnet hysteresis motor (IPMHS) is a self-starting solid rotor hybrid synchronous motor. Its rotor has a cylindrical ring made of composite material like 36% cobalt steel alloy, special Al-Ni-Co, Vicalloy, etc. with high degree of hysteresis energy per unit volume [1-14]. The rare earth permanent magnets are buried inside the hysteresis ring and the ring is supported by a sleeve made of non-magnetic materials like Aluminum which forces the flux to travel circumferentially inside the rotor ring [3, 5-11]. Fig. 1 shows the rotor of an IPMHS motor depicting the position and orientation of permanent magnets. The inclusion of permanent magnets creates rotor saliency without changing the length of the physical airgap and acts an additional permanent source of excitation in the rotor. The induced magnetization of the hysteresis material inside the rotor ring always lags behind the time varying magnetic field. This time lag produces a torque, called hysteresis torque. Also, the induced eddy currents in the hysteresis ring generate some additional starting torque called the eddy current torque"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001892_gt2015-42580-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001892_gt2015-42580-Figure1-1.png",
        "caption": "FIGURE 1. Definition of the axial inclination and other parameter of the brush seal",
        "texts": [
            " Driven by efficiency request, it took some time until brush seals found their way from the aero engines applications into stationary turbo machines for power generation [2,3]. Nowadays, more and more brush seals are used as sealing devices in service (retrofit) applications as well as in new units. In recent years a high experience from brush seal applications in steam turbines has been gained [4,5], and it is obvious that the design standard for steam turbine sealing is different compared to aero engines. With the use of the cold air test facility the axial inclination \u03d5 , see Fig. 1 as a new design parameter as well as the gapping effect for a high clamping of the bristle element between the front and back plate was found and presented [6,7]. The idea to support the bristle pack motions and their adaptiveness with a modified back plate design, to provide brush seals suitable for the respective operating parameters, is widely known and investigated [8,9]. Furthermore, results depending on the bristle pack design e.g. the fence height (c f h) or the initial clearance of the bristle pack to the shaft (cg), show a limited potential to decrease the leakage flow of brush seals [10]",
            " 1) was found which can be also seen in Fig. 4 [7]. The same similar behavior was found for the variation 2 for the seal no. 2, an increased leakage for a high distance reduction. The other variations especially the variation 3 show a special behavior of the bristle pack and will be discussed later in the paper. To calculate the possible bypass through the back plate rings (Fig. 3), the mass flow measurements of the reference back plate were compared to the measurements of the same test seals using a continuous back plate as shown in Fig. 1. A bypass through the back plate rings should increase the mass flow. For the measurements shown here no increasing mass flow was detectable. Thus a bypass leakage effect can be neglected. The blow-down capability of the brush seals was analyzed using photos from a digital microscope looking at the clearanc between the shaft disc and the last bristle row at the back plate. These photos were taken for the unpressurized and pressurized 5 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000556_20140824-6-za-1003.01018-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000556_20140824-6-za-1003.01018-Figure5-1.png",
        "caption": "Fig. 5. Possible cases of candidate grasps: (a) non FC candidate grasp, (b) candidate grasp reducing the current quality, (c) feasible candidate grasp.",
        "texts": [
            " G\u2217 i is formed using a candidate point W \u2217 j \u2208 \u2126l C replacing each of the vertices that define the face FQ. 15: i = i+ 1 16: until (G\u2217 i is a FC grasp or j = m). 17: Compute the quality Q\u2217 for the grasp G\u2217 i . 18: until (Q\u2217 > Ql or j = m) 19: Update l = l + 1 and Gl = G\u2217 i 20: until (j = m) 21: return Gl In Steps (17) and (19) the quality Q\u2217 is calculated for each candidate grasp G\u2217 i ; if a candidate grasp fulfills Q\u2217 > Ql it becomes the new grasp Gl. The procedure continues until no improvement in the quality is achieved. Figure 5 illustrates the three possible cases regarding a candidates grasp; case (a) is an unreliable grasp since it loses the FC condition; case (b) is discarded because the quality of the candidate grasp is smaller than the quality of the current grasp, and case (c) is a candidate grasp that improves the quality of the current grasp, therefore it becomes the current grasp for the next iteration. In this section the proposed optimization strategy is illustrated with some examples for articulated objects with 2 and 3 links"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure9-1.png",
        "caption": "Fig. 9. (a) Magnetic field lines (Wb/m), and (b) magnetic flux density (T) norm of the permanent magnet motor. The range of contour lines is in brackects.",
        "texts": [],
        "surrounding_texts": [
            "linear system (24)] took less than 4% of the solution time of the276 motor magnetic field [assembly and solution of the nonlinear277 DAE (20)].278\nThe difference between B and \u2211\nk Bk should theoretically be279 zero. It was computed for FEM approximations, and the relative280 two-norm281 \u221a \u221a \u221a \u221a \u222b \u03a9 \u2016B \u2212 \u2211 k Bk\u20162 da\n\u222b \u03a9 \u2016B\u20162 da\n(27)\nremained below 5.3e-10, and the difference between the com-282\nputed A and \u2211\nk Ak remained below 7.5e-12 Wb/m for the283 studied time span.284 The stator field components are necessary for calculating the285 torque exerted on the rotor. In that case, applying (11) to the286 rotor iron and bars results in a total of four torque components:287 stator iron to rotor iron, stator iron to rotor current, stator current288 to rotor iron, and stator current to rotor current.289 Fig. 6 presents the total torque on the rotor computed with290 the MST and the sum of all torque components on the rotor.291 In this example, there is approximately a 1.5% difference be-292 tween the computed total rms torques. We have observed that293 the difference decreases with a refined mesh. This is consistent294 with the results presented in [20], where forces obtained with295 the MST and ESM (using external field) are compared with dif-296 ferent mesh densities (the sum of stator field components is the297 external field for the rotor).298 It is evident from Fig. 7 that a major part of the torque ripple299 produced by the squirrel-cage induction motor is due to stator300 iron to rotor iron interaction. In this example, there is also a301\nnotable phase shift between the stator iron to rotor iron torque 302 component and the other torque components. 303\nB. Permanent Magnet Motor 304\nFig. 8 presents the model geometry of the permanent magnet 305 motor and Table II the essential motor parameters. The rotor has 306 four surface-mounted magnets with 1-T remanence flux density 307 and 1.05 \u03bc0 permeability. The BH-curve of the iron cores and 308 shaft are presented in Fig. 3. The iron cores are assumed non- 309 conducting, whereas the shaft and magnets have a conductivity 310 of 4.3e6 S/m and 6.7e5 S/m, respectively. 311\nThe permanent magnet motor fields are decomposed into four 312 components: the stator iron field, the stator coil current field, the 313 rotor iron field, and the rotor magnets field. The magnetic field 314 and its components are presented in Figs. 9 and 10, respectively. 315\nThe relative two-norm (27) was used to measure the differ- 316\nence between B and \u2211\nk Bk . This remained below 4.8e-10, and 317\nthe absolute difference between A and \u2211\nk Ak remained below 318\n7.9e-12 Wb/m for the studied time span. For comparison, in [8], 319 the FR method was applied to a permanent magnet motor, and 320 there was approximately a 5% error in the magnetic flux density 321 in the motor air gap. 322",
            "Fig. 11 compares the total torques exerted on the rotor calcu-323 lated with the MST and with field decomposition. There is ap-324 proximately a 0.2% difference in the computed total rms torques.325 Fig. 12 presents the torques exerted by the stator iron and sta-326 tor coil current on the rotor iron and rotor magnets, respectively.327 It is evident from Fig. 12 that most of the permanent magnet328 motor torque and torque ripple is produced by stator iron to rotor329 magnets interaction.330\nVI. DISCUSSION331\nIn the examples in Section V, the motor magnetic fields were332 decomposed according to iron cores, current-carrying conduc-333\ntors, and permanent magnets. It is possible to decompose the 334 magnetic field in many other ways as well. For example, one 335 could refine, if necessary, the field generated by the stator core 336 into the field components of stator teeth and the rest of the sta- 337 tor iron. Furthermore, the field caused by stator coils can be 338 divided into field components caused by separate phases. The 339 refinement of field components also results in a number of new 340 torque components. 341\nThe results give new quantitative information on the parts of 342 the motor that produce torque. Such insights are useful in, for 343 example, minimizing the torque ripple. For example, the effect 344 of changes made to the shape of the stator tooth, the perma- 345 nent magnet, the squirrel-cage bar, winding distribution, and 346 the feeding current waveform can be seen in the corresponding 347",
            "torque components. Observing the sensitivity of the torque com-348 ponents to these changes can help the designer make more349 sophisticated decisions in optimizing parameter variation; for350 example, she can reduce the dimension of the feasible set.351\nVII. CONCLUSION352\nThis paper has presented a systematic way to isolate and353 quantify electromagnetic forces between different parts of an354 electric motor. Our approach is based on decomposing the mag-355 netic field to so-called magnetic field components generated by356 the magnetization and current density of the different motor357 parts. The forces and torques exerted by these magnetic field358 components were then computed with the conventional ESM359 force and torque expressions.360 The proposed analysis was applied to and demonstrated with361 two typical electric motors: a squirrel-cage induction motor and362 a permanent magnet synchronous motor. These two examples363 confirm that the proposed method can be successfully applied to364 different types of electric motors with nonlinear materials and365 eddy currents.366 Because the distinct interacting parts of an electromechani-367 cal device can be chosen freely, the proposed method offers a368 versatile tool to isolate and analyze force interactions in other369 devices as well.370\nAPPENDIX A371 CIRCUIT COUPLING372\nLet Q denote the index set of stator coils. Then, a stator coil373 with an index q \u2208 Q has a cross section \u03a9q consisting of coil374\nsides \u03a9+ q1 , . . . ,\u03a9+ qn and \u03a9\u2212 q1 , . . . ,\u03a9\u2212 qn . The superscripts + and375 \u2212 denote the orientation of the coil side with respect to the axial376 direction.377 To deal with the number of the turns and the direction of the378 current in the coils, let the source current density Js in the q:th379 coil cross section be380\nJs = Iq\u03b2qez (28)\nwhere Iq is the current in the q:th coil conductor, and \u03b2q is an 381 auxiliary function 382\n\u03b2q (p) =\n\u23a7 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8\n\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9\n+ Nqi\nSqi\nif p \u2208 \u03a9+ qi\n\u2212Nqi\nSqi\nif p \u2208 \u03a9\u2212 qi\n0 if p /\u2208 \u03a9q\n(29)\nwhere Nqi is the number of turns in the i:th coil side \u03a9qi , and 383 Sqi is the cross-sectional area of the i:th coil side \u03a9qi . 384\nThe voltage Vq and current Iq of the q:th coil are related by 385 equation 386\nVq = \u2202t\u03c8q + RqIq (30)\nwhere Rq is the series resistance of the coil and \u03c8q is the flux 387 linkage of the coil in the sense of 388\n\u03c8q = \u222b\n\u03a9q\nleA\u03b2q da (31)\nwhere le is the effective length of the motor. 389\nBecause this paper is focused on the decomposition of the 390 magnetic field and the computation of torque components, we 391 have chosen not to take into account the end effects of the motor 392 as external circuits. Coil end-winding and rotor end-ring circuit 393 models can be included if so desired; for example, see [21]. 394\nAPPENDIX B 395\nTABLE B1 CORE-IRON AND SHAFT-STEEL BH -PAIRS\nH (kA/m) B (T) H (kA/m) B (T)\n0.000 0.000 0.000 0.000 0.079 0.640 0.325 0.671 0.135 0.920 0.459 0.949 0.159 1.010 0.693 1.162 0.190 1.100 1.115 1.342 0.239 1.200 2.137 1.500 0.318 1.300 4.610 1.643 0.493 1.400 9.385 1.775 0.645 1.450 15.099 1.897 0.875 1.500 24.065 2.012 1.273 1.550 35.430 2.121 1.591 1.575 82.402 2.225 2.149 1.600 168.296 2.324 3.342 1.650 237.147 2.419 4.775 1.700 314.601 2.510 6.525 1.750 318.528 2.598 9.151 1.800 451.355 2.683 11.937 1.850 515.834 2.766 15.120 1.900 580.395 2.846 18.542 1.950 642.002 2.924 22.282 2.000 702.720 3.000 27.454 2.050 761.512 3.074 35.810 2.100 819.172 3.146 47.747 2.150 875.384 3.217 63.662 2.200 930.493 3.286 93.901 2.250 984.396 3.354\n(a) Core iron (b) Shaft steel"
        ]
    },
    {
        "image_filename": "designv11_64_0001144_15325008.2014.913737-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001144_15325008.2014.913737-Figure1-1.png",
        "caption": "FIGURE 1. Sketch of five-phase IM winding distribution.",
        "texts": [
            " The third harmonic current is obtained to maintain a quasi-trapezoidal airgap flux distribution and improve the electromagnetic torque density. The proposed SMO with stator resistance adaption online is adopted, and it is utilized to improve the performance for low-speed estimation. Finally, experiments have been performed to validate the effectiveness of sensorless vector control under fault condition. This study adopts a four-pole five-phase IM with identical quasi-concentrated phase windings, which are evenly distributed with a 72\u25e6 spatial angle. Each phase occupies eight stator slots, as shown in Figure 1. To obtain a decoupled model, it is shown that transformation matrices are required in the healthy conditions [5]; obviously, it should be modified when one phase is open. The electromechanical energy conversion takes place in the \u03b11\u03b21 and \u03b13\u03b23 subspaces [7], defined as follows:\u23a7\u23aa\u23aa\u23aa\u23a8\u23aa\u23aa\u23aa\u23a9 C1(0) = [ \u03b11 \u03b21 ] = 2 5 [ cos \u03b3 cos 2\u03b3 cos 3\u03b3 cos 4\u03b3 sin \u03b3 sin 2\u03b3 sin 3\u03b3 sin 4\u03b3 ] C3(0) = [ \u03b13 \u03b23 ] = 2 5 [ cos 3\u03b3 cos 6\u03b3 cos 9\u03b3 cos 12\u03b3 sin 3\u03b3 sin 6\u03b3 sin 9\u03b3 sin 12\u03b3 ] . (1) The stationary transformation matrices T 1(0) and T 3(0) are derived as T 1(0) = [ C1(0) null [C1(0)] ] , T 3(0) = [ C3(0) null [C3(0)] ] , (2) where \u03b3 = 2\u03c0/5, and \u201cnull\u201d is the MATLAB (The MathWorks, Natick, Massachusetts, USA) function for obtaining orthogonal basis vectors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001044_acc.2014.6859161-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001044_acc.2014.6859161-Figure1-1.png",
        "caption": "Figure 1. TORA system configuration.",
        "texts": [
            " In this paper, passivity-based controllers proposed in [1] and [3] are employed to complete the state feedback control design. To design the nonlinear observer a suitable global change of coordinates is defined and the actuated subsystem of TORA is transformed to a flat mechanical system. Then, the nonlinear observer is designed for the TORA system in the new partially flat form. Finally, simulation results are provided to demonstrate the effectiveness of the proposed controllers. The TORA shown in Figure 1 is a simple mechanical system. We let be the normalized displacement of the platform from the equilibrium position and the angle of the rotor. Lagrangian of the simple mechanical systems is in the form of difference between a positive semi-definite kinetic energy and a potential energy [14]. Then, the Lagrangian of TORA can be written as    where is the position vector, is the kinetic energy, is the potential energy of the system and denotes symmetric positive definite inertia matrix. These can be defined for the normalized TORA as [2]      978-1-4799-3274-0/$31"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.29-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.29-1.png",
        "caption": "Fig. 3.29 Double wishbone suspension\u2014kinematic structure",
        "texts": [
            " Having a separate glance on the spatial four-link absolute kinematics Fig. 3.27 Module for absolute kinematics of a complex multibody system mechanism, one can recognize that this loop possesses two degrees of freedom, one for the spring deflection b1 of the wheel mount and a second independent DoF, which corresponds to an isolated rotation of the wheel carrier about the connecting straight line between upper and lower spherical links of the wishbone, which is required for the guidance of the steering motion over the steering mechanism referred to in Fig. 3.29. In a similar manner, the second kinematic loop L2 of the double wishbone axle can be prepared and successively interrelated with the loop L1 over the linear equations in the variables b2; b3; b4 of the spherical joint S1 (Figs. 3.29 and 3.30), corresponding to the coupling of the loops via these joints, to the already shown block diagram for the calculation of the relative kinematics (Fig. 3.31). An important advantage of this approach, which can be applied in its basic ideas to a large number of further technically interesting kinematic structures, is the minimized kinematic calculation amount: Only those kinematic variables must be calculated, which are necessary for the further continuation of the calculation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.2-1.png",
        "caption": "Fig. 3.2 Open kinematic chain and closed kinematic chain",
        "texts": [
            " Choosing one body in the tree as a reference body, one gets nG \u00bc nB \u00f03:1\u00de with nG number of joints nB number of bodies (not including the reference body) D. Schramm et al., Vehicle Dynamics, DOI: 10.1007/978-3-540-36045-2_3, Springer-Verlag Berlin Heidelberg 2014 43 Closed kinematic chains\u2014kinematic loops Assuming an open kinematic chain with tree structure, a single independent kinematic loop is obtained by introducing an additional joint in each case. Using Eq. (3.1), the number of independent loops nL in such a kinematic structure is given by (Fig. 3.2) nL \u00bc nG nB: \u00f03:2\u00de Furthermore, partially and completely closed kinematic chains can be differentiated. A system with kinematic loops forms a partially closed kinematic chain, when \u2022 single partial systems form open chains or \u2022 multiple closed partial systems are connected to each other in an open chain (Fig. 3.3). A chain can be considered completely closed when \u2022 each body is a part of a multibody loop and \u2022 each loop has at least one body that is connected to another loop. A mechanism, by definition, must be a partially or completely closed kinematic chain (for more details see Sect"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001135_j.jappmathmech.2014.09.003-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001135_j.jappmathmech.2014.09.003-Figure1-1.png",
        "caption": "Fig. 1.",
        "texts": [
            "6,8 Three-link systems with a star-like and sequential joining of the links can be brought into an arbitrary osition on a plane.7 However, by virtue of the kinematic constraints between the vertices of three-link systems, their motions are very omplex. The quasistatic motions of a system of three points on which no constraints are imposed are considered in this paper. The interaction orces of the points are the control forces. . Statement of the problem Consider three particles Mi, the masses of which are equal to mi, located on a horizontal plane (Fig. 1) (from now on, unless othrwise stated, i, j = 1, 2, 3). The forces of dry friction act between the points and the plane. For any pair of subscripts i /= j, a force fij, Prikl. Mat. Mekh., Vol. 78, No. 3, pp. 316\u2013327, 2014. \u2217 Corresponding author. E-mail address: t figurina@mail.ru (T.Yu. Figurina). ttp://dx.doi.org/10.1016/j.jappmathmech.2014.09.003 021-8928/\u00a9 2014 Elsevier Ltd. All rights reserved. d r H c o c W m R o i f t irected along the line MiMj, acts from the point Mi onto the point Mj and fij = \u2212fji"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure17-1.png",
        "caption": "Fig. 17. Fixed-anchor Shoe and its True Floating Equivalent",
        "texts": [
            " The main point, however, concerned the very premise on which comparisons were based. For that purpose the author had selected three quantities: shoe factor, spragging, and rate of change in shoe factor (which might be indicative of fade). Now, in practice, the performance of a brake was judged not only by pedal load (shoe factor), but also by pedal travel. The ratio of shoe-tip-travel/shoe-centre-lift was an adequate geometrical yardstick of pedal travel, but had unfortunately been ignored. Had it been taken into account, the problem might have been formulated as on Fig. 17, which showed a fixed-anchor shoe and its true floating equivalent, both having the same shoe-up-travel/ shoe-centre-lift ratio. If then the shoe factors were plotted for the brake illustrated in Fig. 17, and its sliding shoe equivalent, the results shown in Fig. 18 would be obtained. (On the graph both p and the shoe factor were to an inverse scale which resulted in convenient straight-line graphs.) The slope of the sliding shoe factor was some 6 per cent less, that being the approximate difference in fade. That result was no coincidence. A strict mathematical proof could be given to prove the superiority of floating shoes on the premises stated. G SHOE BRAKES FOR ROAD VEHICLES The last basis of comparison might be mathematically correct, but it did not take full advantage of the scope offered by floating shoes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.15-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.15-1.png",
        "caption": "Fig. 6.15 Injection molding cycle",
        "texts": [
            "14) can produce tens of thousands of parts successively in a short period, leading to an effective production rate. Even though the initial cost of the process is very expensive, especially the molding dies, the per-part cost is extremely low since production cost tends to drop greatly as more parts are created. Besides, injection molding produces lower scrap amounts than older manufacturing processes like machining. There are four basic components of the mold that are important in injection modeling: sprue, gates, runners, and vents. Typically, an injection molding cycle consists of four stages (Fig. 6.15): 2. Injection: The raw plastic is first melted by heat and pressure. Then, the molten plastic is forced into the cavity of the mold quickly to ensure no solidification occurs during injection. Two types of machines are used to melt and inject the resin into the mold: a plunger injection molding machine and a reciprocating screw injection molding machine. (Basic design of a reciprocating screw is asked and discussed in Problem 2.7.) 3. Cooling: The molten plastic inside is allowed to cool down and solidify into the desired final shape"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001967_000392184-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001967_000392184-Figure3-1.png",
        "caption": "Fig. 3. Supporting forces, measured at the feet of an upper-leg amputee. Note: The left side of this figure represents the forces measured at the sound leg. The right side represents the forces measured at the artificial leg.",
        "texts": [],
        "surrounding_texts": [
            "Supporting Forces Related to Overload Damage of Leg Amputees\n267\nFigure 4 shows 5 maxima and minima A-E compared with the body weights (= 100 0/0) of the 3 examined groups. The hatching indicates the scatter. Figure 5 shows the divergence of the 3 examined groups expressed as a percentage; force values of normal subjects being 1000/0. The hatching marks the results of statistical significance.\nCompared with healthy subjects, when the foot pushes off, the unload ing of the sound leg as well as of the prosthesis is, on an average, 9 to 100/0 smaller. This means a certain increase in the load.\nThe vertical pushing-off force B of the ball of the foot of the prosthe sis leg is 6-7 Ofo smaller than that in healthy subjects. The horizontal\nD ow\nnl oa\nde d\nby :\nU ni\nve rs\nit\u00e9 d\ne P\nar is\n19 3.\n51 .8\n5. 19\n7 -\n1/ 23\n/2 02\n0 10\n:5 5:\n26 A\nM",
            "KRusEIBAUMANN/GRoH\n268\npushing-off force in the walking direction in the case of the amputees is clearly smaller than in healthy people, viz. 20 Ofo less in the case of the sound leg, and 51 and 36% respectively in the prosthesis. The surplus load C of the amputee in the pushing-off phase is only of short duration and always stays below the maximum values (A, B) of normal subjects. The reduction of the pushing-off forces Band E clearly points to a smaller load in the gait of an amputee.\nD ow\nnl oa\nde d\nby :\nU ni\nve rs\nit\u00e9 d\ne P\nar is\n19 3.\n51 .8\n5. 19\n7 -\n1/ 23\n/2 02\n0 10\n:5 5:\n26 A\nM",
            "Supporting Forces Related to Overload Damage of Leg Amputees 269\nD ow\nnl oa\nde d\nby :\nU ni\nve rs\nit\u00e9 d\ne P\nar is\n19 3.\n51 .8\n5. 19\n7 -\n1/ 23\n/2 02\n0 10\n:5 5:\n26 A\nM"
        ]
    },
    {
        "image_filename": "designv11_64_0000700_978-3-319-11271-8_17-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000700_978-3-319-11271-8_17-Figure1-1.png",
        "caption": "Fig. 1. Scheme of the j-th two-wheeled mobile robot Amigobot",
        "texts": [
            " The second section contains description of the WMR, including kinematics, dynamics and conception of the movement in formation. The third section includes basic information about ADP algorithms, the next section presents hierarchical control system, where are detailed description of the navigator, the WMRF control system and the tracking control system. Section five contains results of the numerical test. The last section summarises the article. The WMRF consists of m WMRs, in the theoretical studies there are used models of the two-wheeled mobile robot AmigoBot. The j-th WMR Amigobot, schematically shown in Fig. 1, consists of two driving wheels (1 and 2), a third, free rolling castor wheel (3) and a frame (4), j = 1, . . . ,m. It has eight ultrasonic range finders s1, . . . , s8 for obstacles detection. Angles between axes of ultrasonic range finders and the axis of the frame of Amigobot are equal \u03c91 = 144\u25e6, \u03c92 = 90\u25e6, \u03c93 = 44\u25e6, \u03c94 = 12\u25e6, \u03c95 = \u221212\u25e6, \u03c96 = \u221244\u25e6, \u03c97 = \u221290\u25e6, \u03c98 = \u2212144\u25e6, the range of individual range finder measurements is equal to di, i = 1,. . . ,8, and the maximal range dmx = 4 [m]. Its movement is analysed in the xy plane [5]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002464_978-981-10-1109-2_5-Figure5.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002464_978-981-10-1109-2_5-Figure5.1-1.png",
        "caption": "Fig. 5.1 Four stages of unfolding of a physical model of fPZM. From 1 to 4: \u03c8 = 0 at stowed position, two intermediate positions, and \u03c8 = \u03c0 at fully deployed position. The angle between faces B and T reaches \u03b6 by the completion of deployment (see Sect. 1.1.1)",
        "texts": [
            " Astraightforwardwayof folding aPZMis by taking advantage of its planar symmetry between the bottom (B) and top (T ) faces. The intersection of this symmetry plane and perpendicular trapezoidal facets form the axes of revolution for the folding. Therefore the fold of the module is a function of angle \u03c8 between the halves of the side facets. For each facet the angle \u03c8 is the same. It would be particularity practical, since each panel or sub-group of panels could be folded by synchronized actuators. Figure5.1 shows a physical model of a foldable Pipe-Z module (fPZM) in the \u201coutside-in\u201d folding scheme, where the side elements are \u201cfolded out\u201d for stowage. This model has been fabricated from corrugated board to reflect the rigidity and to some degree, realistic thicknesses of the elements. fPZM is a rigid-panel structure of trapezoidal panels connected by cylindrical hinges. Thus it is vital that the parts of a fPZM are not distorted during folding. Preliminary experimentation with the physical models including the one shown in Fig. 5.1 suggested that fPZM is a proper mechanism. Further geometrical analysis has also shown that, as illustrated in Fig. 5.2. Although the fold is a function of the side angles \u03c8 , it is linked to the angle (\u03b6 \u2217) between faces T and B of a fPZM. Obviously, for \u03c8 equal to 0 and \u03c0 , the values of this corresponding central angle are: 0 and \u03b6 , respectively. Angle \u03c8 does not depend on the number of sides n. The trigonometrical relationships of a folded triangular fPZM (n = 3) are shown in Fig. 5.3. Based on the relationships shown in Fig",
            " The same folding scheme, however, based on inflatability could avoid the last two problems relatively easily. Nevertheless, reaching proper rigidity of an elongatedmodular construction (without internal reinforcement) seems very difficult, if possible at all. Nonetheless, CTR system is neutral to under- and super-pressure. A CTR structure composed of a relatively few inflatable units seems rather feasible. Packing capability of CTR is also good (see Fig. 5.8). A low-fidelity six-unit octagonal dPZ has been fabricated. Similarly to the physical model of the hexagonal fPZM (see Fig. 5.1), the panels have been made of thick corrugated cardboard. For easy identification, the fPZMs have been made in contrasting colors. The units are connected by internal tubular elements. This connection gives one degree of freedom (1DOF), (rotation) between every two adjacent units. Each fPZM has two discrete states: stowed and erected. The transitions between these states are done manually. Deployed units are \u201crigidized\u201d by external band clamps made of rubber bands. Figure5.9 shows the deployment of a helical dPZ"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000387_978-1-84996-080-9_7-Figure7.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000387_978-1-84996-080-9_7-Figure7.11-1.png",
        "caption": "Fig. 7.11 Model (a) and real vehicle (b) with surface operation buoyancy configuration \u2013 positively buoyant",
        "texts": [
            "8, used to visualise a VR scene2 in real-time, using measurements from virtual sensors (simulation environment) or real sensors (real-world environment). Figure 7.9 displays a part of the main screen of the ROV simulator. A part of the PC104 display is shown in Figure 7.10. This display is not available in the realworld environment. However, the PC104 can be accessed and remotely administered from the Control PC through the network using the remote administrator. 2 Ship, ROV, ocean surface, seabed and so forth The Thrusted Pontoon/ROV is a multi-mode of operation vehicle. It can be operated on the surface (Figure 7.11) as a survey platform either towed (by an extra towing cable) or thrusted by 4 horizontal thrusters to allow surge, sway and yaw which is useful in confined spaces or near hazards where a boat and tow cannot operate. It can also be operated as an ROV (Figure 7.12). In these various modes of operation, it is used in conjunction with an umbilical and associated winch; the umbilical carrying vehicle power, control and data from sensors and instruments. In the surface-tow or surface thrusted modes of operation, overall vehicle buoyancy is maintained strongly positive by 8 buoyancy modules mounted on the pontoon upper frame"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure12-1.png",
        "caption": "Figure 12. Switch-off mechanism.",
        "texts": [
            " By the rotation of Plate C, Pin C switches off and each Claw C moves along each Guide Hole C (Fig. 10(d)). After that, four Claws C simultaneously mesh with Ratchet Wheel C and thus Shaft C is locked. The lock of Shaft C is released by rotating Shaft C in a direction opposite to the direction locked by the safety device. The torque-based safety device consists of a commercial torque limiter (Fig.11, TGB20-HC, Tsubakimoto Chain Co., see [14] for more details) and a switch (Switch Z). As shown in Fig. 12, Switch Z is installed at the position of being pressed by Plate Z of the torque limiter when Plate Z slides along Hub of the torque limiter. Switch Z can interrupt electric power supply to all motors of the robot. Shaft A is connected to Hub. Hub has a threaded portion and Adjusting Nut is attached to Hub. Balls are respectively inserted into Holes of Center Flange and set in V-shaped Pockets of Hub. Balls are held in V-shaped Pockets by the spring force generated by Disk Springs. Center Flange is attached to Parts A. Shaft B is connected to Parts A. When the magnitude of the Shaft A\u2019s torque does not exceed the detection torque level, the torque is transmitted to Shaft B via Hub, Balls, Center Flange and Parts A. If the magnitude of the torque exceeds the detection torque level, Balls pop out of V-shaped Pockets and Plate Z slides along Hub and switches off (Fig.12). The detection torque level is adjustable by changing the attachment position of Adjusting Nut. We developed the walking support robot equipped with four velocity-based mechanical safety devices and two torque-based mechanical safety devices. Fig. 13 shows the developed robot. As shown in Fig. 13, the length is 125[cm] and the width is 154[cm]. The armrest is adjustable in height from 85[cm] to 108[cm] according to the height of each patient by using a hand crank. We conducted the following experiments by using the developed walking support robot"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001003_s40314-014-0146-7-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001003_s40314-014-0146-7-Figure3-1.png",
        "caption": "Fig. 3 Illustration of control thrusters arrangement. This scheme repeats in the other four sides",
        "texts": [
            " Finally, one shall note that the elements of are related with the convergence rate of the parameter estimation law. Regarding chattering, in the same way of the standard controller, the sign function should be changed by a saturation function in order to mitigate its effects. 4.1 Thrusters arrangement and technology The control design of Sect. 3 assumes that a pure torque can be generated by the attitude control thrusters without inducing translation, this effect can be obtained by using a configuration of anti-symmetric thrusters. Figure 3 illustrates one of the many possible configurations that generate the desired torque with no translation. Naturally, this scheme shall be repeated for all the 3 axis of rotation, resulting in 12 control thrusters. Note that a thruster, in fact, generate a force, the torque results from the respective arm until the spacecraft\u2019s cm. In practice, the control torques can assume only three levels: zero, minimum negative and maximum positive. The force, in fact, can be only zero or positive along the thruster axis; for a specific control axis, positive torque is generated by choosing one pair of thrusters, negative torque is generated by the other pair"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001615_1.4031579-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001615_1.4031579-Figure10-1.png",
        "caption": "Fig. 10 Slow speed stiffness test facility cross-sectional schematic",
        "texts": [],
        "surrounding_texts": [
            "Key parameters defining the two prototype seals and the coverplate configurations (see Fig. 16 for legend) used during the tests are given in Table 4. The leaf packs are similar; however, the Table 1 Oxford slow-speed stiffness test facility capability and measuring ranges Property Minimum Maximum Upstream pressure 17 kPa 1250 kPa Downstream pressure Ambient Eccentricity range 2.5 mm \u00fe 2.5 mm Eccentricity speed 0.008 mm s 1 0.36 mm s 1 Shaft rotation 60.03 rad s 1 61.37 rad s 1 Torque n/a 650 N m Eccentric force n/a 6500 N Mass flow rate n/a 1.0 kg s 1 Table 2 95% confidence level uncertainty Measurement Systematic error Random error Time 3 10 4 s 1 10 6 s Force 0.89 N 0.73 N Displacement 9:1 10 3 mm 0.010 mm Torque 0.18 N m 0.44 N m Temperature 2.90 C 0.2 C Pressure (0\u20131600 kPa) 3600 Pa 1700 Pa Pressure (0\u2013500 kPa) 2100 Pa 400 Pa Journal of Turbomachinery JANUARY 2016, Vol. 138 / 011004-7 Downloaded From: http://turbomachinery.asmedigitalcollection.asme.org/ on 01/27/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use coverplate arrangements are significantly different. Prototype A only uses a rear coverplate, thereby resembling the configuration used in the numerical model. Prototype B uses both a front and rear coverplates. Extensive experimental data showing how blowdown and lift-up are linked to coverplate configuration were reported in Ref. [14]. Prototype A exhibits strong blow-down with its tested coverplate configuration, whereas B is expected to experience only slight blow-down or even a balanced operation. For Prototype B, the dynamic friction coefficient between leaf elements and rotor, ld, was measured by Peel [22]. 5.1 Prototype A. A torque characterization test spanning the upstream pressure range 0\u2013400 kPa was carried out. The recorded torque values for both the increasing and decreasing pressure ramps are shown in Fig. 17. This shows that the torque exerted on the rotor by the seal increases with pressure and that the relationship is close to linear. The fact that the torque during the decreasing pressure ramp is slightly higher is a sign of hysteresis in the leaf pack due to interleaf forces and friction. However, the comparatively small magnitude of the hysteresis shows that these interleaf frictional forces do not play a significant role in this leaf seal. The results from the stiffness characterization tests, at the discrete pressure levels between 0 and 400 kPa, are shown in Fig. 18. Seal stiffness is defined as force per unit incursion and is the gradient of the force versus displacement plot. For 0 and 80 kPa a positive stiffness loop is observed. Here the force acting on the rotor acts in a direction which attempts to recenter the seal and rotor. This force increases as the magnitude of the relative displacement increases. The force offset between the two lines at d \u00bc 0 is caused by stiffening of pressure supply pipes in the test facility, which creates a systematic error in measured force. However at pressure differentials of 160 kPa and above, the seal stiffness has decreased and becomes increasingly negative as pressure increase. This implies that the force exerted on the rotor by the Fig. 12 Displacement cycles used to characterize the seal Fig. 13 Eccentric casing incursion cycle, measured by displacement sensor (prototype A, PU 5 200 kPa) Table 4 Test seals and coverplate configurations Dimension Prototype A Prototype B Lay angle (deg) 60 60 Leaf thickness (mm) 0:08 0:08 cU (mm) n/a 1.3/1.8/2.0 cD (mm) 2.2 1.3/1.8/2.0 xU (mm) n/a 0.3 xD (mm) 0.5 0.3 ld \u2014 0:6 011004-8 / Vol. 138, JANUARY 2016 Transactions of the ASME Downloaded From: http://turbomachinery.asmedigitalcollection.asme.org/ on 01/27/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use seal is in such a direction that the rotor is pushed in the direction of increasing eccentricity. The plots at higher pressure differentials are clear examples of negative seal stiffness. During tests, video was recorded at the location of maximum seal incursion. Figure 19 shows a screenshot corresponding to maximum incursion and excursion for the test with an upstream pressure of 200 kPa. In both cases, the leaf tips can be seen to be in contact with the rotor, meaning that they still fulfil their sealing function. This is in agreement with mass flow rate data plotted Journal of Turbomachinery JANUARY 2016, Vol. 138 / 011004-9 Downloaded From: http://turbomachinery.asmedigitalcollection.asme.org/ on 01/27/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use against rotor eccentricity recorded at the same condition and shown in Fig. 20. Mass flow rate is seen to increase by less than 2.5% over the entire movement range. In contrast, the formation of a gap between the leaf tips and the rotor would have created an increase in flow area and measurable change in mass flow rate. 5.2 Prototype B. Prototype B was subjected to the same characterization tests; however, the tests were carried out using the three different test rotors listed in Table 3. The resulting data are plotted in Fig. 21. In the case of matched seal bore and rotor radius (R0), the torque initially increases linearly up to a pressure differential of 150 kPa, at which point there is a reduction in gradient followed by a further linear trend. The rapid increase in gradient at low pressures is attributed to the seal bore not being perfectly round, implying that during this initial phase the number of leaves in contact with the rotor increases with pressure. In the second phase, contact exists around the entire circumference and the true torque characteristic for the seal is observed. This hypothesis is confirmed by the large rotor data (R0 \u00fe 0:5). Here due to the interference, all leaves are in contact from the start and consequently a linear relationship is observed over the entire pressure range. Finally for the smallest rotor (R0 0:2), again two regions are observed. At low pressures, the torque increases rapidly, which can be attributed to more leaves coming into contact with the rotor; however at higher pressure differentials, the torque plateaus. Prototype B exhibits a stiffness characteristic typical of contacting seals. As can be seen from Fig. 22 for all three rotors, seal stiffness increases with increasing pressure differential being applied to the seal. For the small and nominal rotor sizes, the seal stiffness is seen to double between the 0 kPa and 100 kPa test. This is due to blow-down bringing the entire seal in contact with the rotor, thereby doubling the seal stiffness. This effect is in agreement with Eqs. (29) and (21) developed earlier. The same is not true for the large rotor case. Here due to the interference between the seal and rotor, all leaves are in contact with the rotor from the outset and will remain so, even during excursions. For none of the tests, negative stiffness was observed. However, when looking at the trend of mean stiffness (corresponding to excursions with 0.4< d< 0.4 mm, in Fig. 22) against pressure drop, shown in Fig. 23, all three tests show a gradient change at 100 kPa. In the nominal configuration, the gradient becomes negative in the region 100\u2013300 kPa, before continuing to increase at the same gradient as the large and small diameter configurations. To some extent, the lower impact of the aerodynamic forces can be explained by the fact that prototype B is significantly stiffer (Fig. 23). However, it is expected that some of this change is also caused by the different coverplate geometry, which will influence the seal operation. Mass flow rate against displacement data were also recorded for all prototype B configurations. Flat characteristics, similar to the one shown in Fig. 20, were observed. The variations were less than 7%, 4%, and 2.5% for the small, nominal, and large rotors, respectively. 5.3 Comparison of Prototypes A and B. Based on the experimental data collected and presented in Figs. 17\u201323, the following conclusions can be drawn with respect to the seal: \u2014 Both seals exhibit a close-to-linear increase in torque as pressure is increased. The exceptions to this are the small and nominal diameter rotor cases of prototype B, for which at low pressures only a fraction of the seal circumference is in contact with the rotor. \u2014 Prototype B shows a slightly larger increase in torque with pressure, suggesting that aerodynamic blow-down is somewhat stronger in the configuration with two coverplates. \u2014 Prototype B is mechanically stiffer. Comparing stiffness at DP \u00bc 0 for the two nominal rotor cases shows an almost sixfold stiffness increase. \u2014 Prototype A shows a continuous stiffness reduction as pressure differential is increased. In contrast, prototype B only exhibits stiffness reductions for the nominal diameter seal at intermediate pressures (50\u2013300 kPa) and the large diameter seal at high pressures (>500 kPa). For the small rotor, this can be explained by the simultaneous change of pressure and number of circumferential sealing elements in contact with the rotor (due to blow-down). In the case of the large rotor, the likely cause is the significant compression of the leaves (> 0.5 mm) resulting in a nonlinear mechanical force that outweighs the aerodynamics contributions until large pressure drops. \u2014 For all the tested seals, the leakage performance is independent from displacement d. The largest variation was 67%, but 62.5% was more typical."
        ]
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure6-1.png",
        "caption": "Fig. 6. Deformation of the model, case #3",
        "texts": [],
        "surrounding_texts": [
            "One of the most important questions regarding to the mechanical loading conditions of any structure is the distribution and the maximal value of forces affecting to the different nodes. Fig. 8 shows the distribution of the forces in the model. The maximal force value was 611.48 N(y) in case #1, 754.86 N(y) in case #2 and 1718.55 N(y) in case #3. It can be determined that asymmetry has a notable effect on the magnitude of maximal forces affecting to the structure. In case of symmetric loading conditions, none of the partial forces reached value of the total mechanical load (1500 N)."
        ]
    },
    {
        "image_filename": "designv11_64_0003742_9781119260479.ch9-Figure9.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003742_9781119260479.ch9-Figure9.1-1.png",
        "caption": "Figure 9.1 The low-inductance rotor configuration of a PMSM. The image on the left illustrates a nonsalient pole configuration. In the right image, the steel rim is made as thin as possible on the d-axes to reduce machine inertia. The flux path at the d-axis is sufficient for the excitation flux (see the magnetic flux lines at the d-axis), and magnetically the machine is slightly asymmetric, since there is higher reluctance on the quadrature axis than on the direct axis (see the flux lines at the q-axis). The q-axis path on the right shows higher reluctance than the d-axis path. However, because of the large magnetic air gap, Ld Lq, which is typical of rotor surface magnet PMSMs.",
        "texts": [
            " Several of the introduced methods are based on the use of computationally intensive estimators. Good results have been achieved in several studies for operation at moderate supply frequencies. However, operating at close to zero speed remains somewhat problematic. 9.1 PMSM configurations and machine parameters The performance characteristics of a PM machine are highly dependent on rotor structure. PMSM rotor poles can be implemented in various ways. The simplest is to mount magnets directly onto the rotor yoke surface. See Figure 9.1. The yoke can be a simple steel tube if the stator harmonics are at sufficiently low frequencies. However, in machines with higherfrequency stator harmonics, it is better to use a laminated structure to minimize the eddy current losses that can lead to excessive rotor temperatures. Rotor surface magnet machines are very low in magnetizing inductance. If they also have a small number of poles, leakage inductance also remains low. Therefore, a voltage source inverter with a high switching frequency is required for best current behaviour"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure13-1.png",
        "caption": "Figure 13 Pulling mechanism (see online version for colours)",
        "texts": [
            " By this mechanism, the centre of rotation of each joint corresponds to the centre of the manipulator. Thus, problem P1 in Section 4 is solved. In addition, the size of the duplex system is reduced from that of the conventional one. The locking mechanism was also improved to realise the semi-circular shape in Figure 12. By pushing the piston attached to the end of the manipulator, the hose is expanded by air, and the pins engage to the holes of the link. This locks the joint. To solve P2, a pulling mechanism was employed. Figure 13 shows the pulling mechanism. Both manipulators have the same mechanism, but only one is shown in order to simplify the illustration. As shown in Figure 14, by pulling the wire of the locked manipulator, the movable manipulator is pulled forward. This reduces the friction, and the duplex manipulator can escape from its deadlock. By this mechanism, we can utilise both pushing and pulling operations. The pushing operation is effective for the outside manipulator, and the pulling operation is effective for the inside manipulator"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002778_j.proeng.2016.07.095-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002778_j.proeng.2016.07.095-Figure5-1.png",
        "caption": "Fig. 5. Assembly of the helical-gear set: (a) layout sketch;(b) mating of teeth in the pitch point; (c) 3D model of the gear",
        "texts": [
            " We extend the tooth contour by the command Swept Boss/Base, having set the parameter Merge Result and indicated the helix We cut the tooth, which goes beyond the limits of the gear, with the help of the command Extruded Cut, having performed the fillet curve of the tooth f = 0,4mn = 1,6 mm; the chamfer 2 \u00d7 45\u00b0 (Fig. 4, c). By the command Circular Sketch Pattern we distribute 41 teeth over the surface of the root cylinder, by so doing we finish construction of 3D model of the gear (Fig. 4, d). To construct the wheel rim we need to repeat the same operations as for the gear, but with the parameters for the wheel. We open a new document 3D arrangement of parts and/or other assemblies. We activate the command Layout at the panel Layout and make a layout sketch (Fig. 5, a). We specify the pitch point P in the initial point and draw from it two vertical segments, which are equal to the reference diameters of the gear and the wheel. Though the bisecting point of the segment we draw the gear and wheel axes, the length of which is equal to their width. Later on we carry out the following actions. Assembly>Insert Components in the appeared window Property Manager we select the gear file. For the appeared on the screen gear model we specify mating: Concentric with the axis, Coincident of the lateral face of the rim with the end point of the axis and we apply the same operations to the wheel model. For matching the gear teeth to the wheel space we specify the mating Coincident of two adjacent faces of teeth with the pitch point (Fig. 5, b). As a result, we have a 3D model of the helical gear (Fig. 5, c). To check the accuracy of the constructed model we conduct a number of studies [13, 14]. In the normal section deviation of length of the constant chord and the tooth depth from the calculated data didn\u2019t exceed ~5 m of the pressure angle \u2013 0,033% (Fig. 7). The base tangent which characterizes distribution and a value of clearance between teeth in the gear has been calculated according to the following equation [12]: cos [ ( 0,5) 2 ] 4 0,9397 0,5 2 0 0,0156 ,W m z x tg z inv z tg z n w t w zw \u2013 number of teeth included in the base tangent; x \u2013 shifting coefficient, in our case 0;x t t tinv tg \u2013 involute function"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003780_intmag.2015.7157122-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003780_intmag.2015.7157122-Figure1-1.png",
        "caption": "Fig. 1 (a) Exploded diagram of the proposed machine. (b) Magnetic field distribution under normal operation. (c) Magnetic field distribution under faulty operation.",
        "texts": [
            " This paper investigates the fault signatures of a magnetless FSDC generator under armature winding faults . It focuses on the study of short circuit faults and open circuit faults . The prototyped machine is operated with 10% to 50% turn shorts in one of phase windings . Open circuit fault is tested in one of the phases . The rectified generator output current and torque are utilized as the fault indicators . Both simulation and experiment results are studied for the verification . II . FAULT SIGNATURE APPROACH The proposed magnetless FSDC machine is implemented by using software JMAG as shown in Figure 1(a) . By using time-stepping finite element method (TS-FEM), the magnetic field distribution of the proposed machine under normal condition and half-phase A short circuit fault are showed in Figure 1(b) and 1(c) respectively . It can be found that the magnetic field is evenly and symmetrically distributed under normal operation, while it turns out to be asymmetrical under faulty operation (especially the circled areas) . In this study, motor current signature analysis (MCSA) is used as the main fault detection method . It aims to sampling the harmonics components in the stator current spectra via fast Fourier transform (FFT) . The corresponding frequency spectrum is given by: fa = [(1-s)* n/p \u00b1 k ]*f1 (1) Where k = 1, 2, 3, 4\u2026, n is the number of rotor bars, p the number of pole pairs, s the motor slip, and f1 the fundamental frequency "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000300_icra.2017.7989440-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000300_icra.2017.7989440-Figure4-1.png",
        "caption": "Fig. 4. Scan conversion in 3-D required to obtain the metric coordinate.",
        "texts": [
            " In our case, we use the implementation of this algorithm provided by the VTK library (Visualization Toolkit [11]) with the function vtkPolyDataConnectivityFilter to obtain the largest stiff region Vsb and the vtkCenterOfMass function to calculate the center of mass (pc = (ic, jc, kc)) of Vsb. We show in Fig. 3 the mesh obtained to compute the center of mass. We need to convert the value of pc in the metric coordinates pg . To do this, as we use a convex ultrasound probe, we perform a scan conversion of each point inside the RF volume to the Cartesian coordinates (see Fig. 4), s(i, j, k)\u2192 p(x, y, z), in order to obtain the metric location with respect to the Cartesian frame. We define the scan conversion using the ultrasound probe parameters as described in [5]. In our case, RF data is considered instead of pre-scan images. We briefly recall the scan conversion formulation as x = r sin\u03c6 (5) y = [r cos\u03c6\u2212 (rp \u2212 rm)] cos \u03b8 + (rp \u2212 rm) (6) z = [r cos\u03c6\u2212 (rp \u2212 rm)] sin \u03b8 (7) where rp and rm are the radii of the ultrasound probe and the motor of the probe respectively (see Fig. 4 right). The coordinates in the Cartesian volume are sorted as quasispherical coordinates with r as the distance from the point to the origin of the scanlines, \u03c6 as the azimuthal angle in the x-y plane and \u03b8 as the zenith angle (see Fig. 4 right). The quasi-spherical coordinates are computed in function of the RF coordinates as r = vs fs j+ rp, \u03c6 = \u22120.5\u03b1l(Nf \u2212 1)+\u03b1li and \u03b8 = \u22120.5\u03b7(Nf \u2212 1 \u2212 2k) where vs is the speed of the sound (1540 m/s), fs is the sampled frequency, \u03b1l is angle between neighboring scanlines and \u03b7 is the angle of the FoV of the motor in the ultrasound probe for a motor angular step. Every slice function of Vr, fk(i, j) \u2208 Vr, is converted to b-mode (brightness mode), fkb (i, j), image as fkb = log{\u2016Hilbert (fk) \u2016} where Hilbert ( ) is the Hilbert transform of ( )"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003857_0954406215589843-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003857_0954406215589843-Figure6-1.png",
        "caption": "Figure 6. Pad surface grid.",
        "texts": [
            " However, the updated pressure distribution still requires conversion to the generalised forces and moments used in the equations of motion (18) and (19). Conversion of air film pressure to forces The pressure distribution on the pad surface results in a force and a moment on the pad Fpres \u00bc Z Z A npad PdA Mpres \u00bc Z Z A n npad PdA \u00f020\u00de where A Surface area of the pad P Pressure on the pad surface minus the ambient pressure P \u2013 Pa at UNIV OF CONNECTICUT on June 4, 2015pic.sagepub.comDownloaded from npad Normal vector on the pad surface, see Figure 6 n Vector to position on the pad surface In the pad body fixed frame, expressions for the position and normal vector of a pad surface point are constants. np is given in equation (1), and the normal vector is xnpad,p ynpad,p znpad,p 2 64 3 75 \u00bc sin p \u00f0 \u00de 0 cos p \u00f0 \u00de 2 64 3 75 \u00f021\u00de The pressure has been calculated using a grid on the pad surface, see Figure 6. The same grid can be used to calculate the pressure force and moment components with respect to the body fixed frame Fpres,p \u00bc X i X j n i,j\u00f0 \u00de pad,p P i,j\u00f0 \u00de A i,j\u00f0 \u00de Mpres,p \u00bc X i X j n i,j\u00f0 \u00de p n i,j\u00f0 \u00de pad,p P i,j\u00f0 \u00de A i,j\u00f0 \u00de \u00f022\u00de Transformation is required to obtain the force and moments associated with the generalised coordinates of the pad: r, and rF \u00bc zF \u00bc xFp yFp zFp M \u00bc xMp \u00fe zMp M \u00bc yMp \u00f023\u00de Besides the pressure there is also friction in the air gap resulting in a force and a moment on the pad"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001797_humanoids.2015.7363511-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001797_humanoids.2015.7363511-Figure4-1.png",
        "caption": "Fig. 4. (a) When a robot is pushed, the high level controller determine whether to trigger the reactive stepping controller or not. (b) In preparatory phase, the robot attempts to move its CoM to the support foot while passively moving to the pushed direction. If an indicator pl leaves the scaled support polygon, the controller proceeds to lifting phase. (c) In lifting phase, the robot lifts the swing foot to some desired position, which is determined from the desired stepping point sdes. The controller updates sdes repeatedly at every control time step during the lifting phase to deal with continuous perturbations. (d) In landing phase, the robot attempts to land the swing foot to sdes.",
        "texts": [
            " In the lifting phase, the robot lifts the swing foot to some target position, and subsequently in the landing phase, the robot attempts to land the swing foot to the target stepping point. Each phase operates sequentially: when a certain condition is satisfied in a phase, the controller makes transition to the next phase. After the landing phase, the higher level controller checks the stability and determines whether to resume the reactive stepping controller for additional stepping or trigger a postural balance controller. The configurations of a humanoid robot\u2019s legs for each phase are illustrated in Fig. 4. The reactive stepping controller is constructed on top of the momentum controller. Hence, the desired control action is achieved by giving suitable inputs to the momentum controller. Similarly to the postural balance control, we use (1) and (2) in order to determine the desired momentum rate changes l\u0307d and k\u0307d. Therefore, the total inputs to the momentum controller are rGd, ld,kd, \u03b8\u0308 u d ,T d, and vd. Throughout the stepping control, kd is set to zero in order to stabilize the rotational motion"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure4-1.png",
        "caption": "Fig. 4. Graphical Method of Predicting the Lining Wear with a Typical Pivoted Shoe",
        "texts": [
            " Other items are defined as follows :- Leading shoes: shoes in which the movement of the Trailing shoes: shoes in which the movement of the * An alphabetical list of references is given in Appendix 11, p. 51. drum over the lining is towards the pivot or abutment. drum over the lining is towards the applied load. Toe and heel: the ends of the lining in relation to the \u201cankle\u201d at the pivot or abutment. Shoe tip : the part of the shoe at which the operating load is applied; that is, the end remote from the pivot or abutment. pivoted Shes . It is easy to predict, by the simple graphical method shown in Fig. 4, how the lining of a pivoted shoe will at UNIV NEBRASKA LIBRARIES on August 25, 2015pad.sagepub.comDownloaded from INTERNAL EXPANDING SHOE wear. With the shoe pivot as centre, an arc is struck passing through the centre of the brake. A circle of radius equal to that of the drum is then drawn from a centre on the arc at a distance from the drum centre equal to the maximum lining wear. The position of the point of maximum wear will lie on the line which passes through the centre of the circle and the drum centre, and which will make with the line through the pivot and the drum centre an angle of 90 deg"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002774_speedam.2016.7525811-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002774_speedam.2016.7525811-Figure3-1.png",
        "caption": "Fig. 3. Proposed Sensing diagrams.",
        "texts": [
            " Due to notch filter characteristic, one signal which has the resonant frequency has the lowest amplitude. Comparing the signals amplitudes with the reference voltages, the lowest amplitude can be detected. Indeed, signal with the lowest amplitude has a frequency that certainly is resonant frequency of the notch filter at all stages of the rotation. This technique is achieved rotor position of the SRM with a number of reference points from aligned to unaligned states. The proposed sensing diagram circuit is illustrated in Fig. 3. The modulator block consists of one PIC microcontroller and one power amplifier. Pseudosinusoidal signals are produced via the digital-to-analog (D/A) microcontroller stage and are applied to the resonant circuit through the power amplifier. In addition, as seen in Fig. 3, a capacitor (C), a resistance (R), and the motor phase (L( )) are used to establish a resonant circuit. In this circuit, capacitor and motor phase acts as a notch filter. The voltage amplitude which is placed on the resistor R is considered as demodulator input. Therefore, according to Fig. 2, from aligned to unaligned positions, the inductance profile of SRM is divided into 2n modes. Therefore, a resonant frequency is obtained for each mode. In (1), the resonant frequency of an SRM for each point of the motor's inductance is achieved"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000816_s10409-015-0016-6-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000816_s10409-015-0016-6-Figure5-1.png",
        "caption": "Fig. 5 Contours of the dimensionless radial velocity on a the horizontal plane \u03c2 = z/a = 1.0 and b the vertical plane \u03b7 = y/a = 0",
        "texts": [
            " Precisely, if \u03c1\u2032 < \u03c1\u20320 (\u03c2 < \u03c20, respectively), V\u03c1\u2032 increases with \u03c1\u2032 (\u03c2, respectively); otherwise, the opposite tendency is observed. From Fig. 4a, V\u03c1\u2032 vanishes at the points on z-axis (\u03c1\u2032 = \u03c1/a = 0), owing to the axisymmetric property inherent to the present problem. It is also seen from Fig. 4b that V\u03c1\u2032 at the point \u03c2 = z/a = 0 is equal to zero due to the satisfaction of the boundary condition specified by Eq. (15). Parallel to the dimensionless pressure, the contours of the dimensionless radial velocity V\u03c1\u2032 are displayed in Fig. 5. Identical to Fig. 3a, the contours of the dimensionless radial velocity, independent of the variable \u03c6, are concentric circles (see Fig. 5a). In contrast to the dimensionless pressure, which is symmetric with respect to the horizontal plane z = 0, the dimensionless radial velocity is antisymmetric with respect to the same plane, as shown in Fig. 5b. Figure 6 plots the dimensionless axial velocity as functions of the dimensionless coordinates. It is seen from Fig. 6 that V\u03c2 generally deceases with both \u03c1\u2032 = \u03c1/a and \u03c2 = z/a. In particular, for \u03c1\u2032 > 1.0, V\u03c2 increases slightly with \u03c2 = z/a in the neighborhood of \u03c2 = 1.0. The velocity with a radial coordinate \u03c1\u2032 1.0 is much larger than that corresponding to \u03c1\u2032 > 1.0. For completeness, the contours for V\u03c2 are shown in Fig. 7. As expected, the contours on the horizontal plane \u03c2 = z/a = 1.0 are a series of concentric circles (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003921_pime_auto_1949_000_010_02-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003921_pime_auto_1949_000_010_02-Figure5-1.png",
        "caption": "Fig. 5. The Geometry of the Sliding Shoe with Parallel Abutments",
        "texts": [
            " Automatic adjustment has, with one exception, been applied only to pivoted shoes, the positive location in the off-position having been considered desirable, if not essential. The exception is significant and automatic adjustment of sliding shoes is likely to be developed in the future. Sliding Shoes. Because a sliding shoe can centre itself in the drum, the centre of the wear pattern is not fixed, as with the pivoted shoe, by the motion about its pivot, but depends upon the coefficient of friction between the lining and the drum (Fig. 5). The lining has to be mounted on the shoe in a position Effect of change of angle of friction on the position of the centre of pressure and point of maximum wear. which anticipates a certain friction value. If this is exceeded, the lining will wear more at the heel if the shoe is leading or at the toe if it is trailing, and vice-oersa if the friction is less than expected. In the G i r k g 2 L.S. design (Fig. 2b) the shoes are completely reversible, and the wear pattern is centralized by inclining the face at the shoe tip relative to that at the abutment, but with the parallel abutments of the Lockheed design (Fig",
            " k = the distance in terms of drum radius from the centre of the brake drum to the pivot, or, in the case of a floating shoe, the perpendicular distance from the centre to the line of action of the abutment force. The assumptions made in the case of the pivoted shoe are that the line through the centre of pressure and the drum centre is at 90 deg. to the line through the pivot and the brake centre, and that the perpendicular distance from the pivot to the line of action of the shoe-tip load is equal to 2k x r, as shown in Fig. 14. . In the case of the sliding shoe, it is assumed that the shoe tip and the abutment forces are equidistant from the centre, as shown in Fig. 5, where OA = OB = h x r . The value of 1 depends on the included angle of the lining and is likely to lie between 1-1 and 1-17 (Barford 1933). The maximum value of k is limited by the necessity to house the pivot or abutment within the brake drum. It is diflicult for it to exceed 0.8, and 0-75 is a typical value. Leading shoes will \u201csprag\u201d or become self-applying when cot 0 or cosec 0 equals Elk, as the case may be. It follows that maximum stability is obtained when the shortest possible lining is used, and when the pivot or abutment is as close as possible to the drum",
            "comDownloaded from INTERNAL EXPANDING SHOE BRAKES FOR ROAD VEHICLES 51 deceleration will be plotted against pedal effort, together with the brake temperatures, if they have been obtained. Throughout their working life, the brake linings and drums of rubbing speeds, unit pressures, and temperatures, which produces ever-changing surface finishes and conditions, and it must be remembered that it is the behaviour of the type of surfaces prevailing at the moment which is being tested. It is the study of surface phenomena and the search for better cooling which form the fundamental basis for progress in the science of braking. In Fig. 5, (2-X) - 5 - A N - L - BN are undergoing repeated surface treatment under a wide range x k-Z sin 0 therefore x = ~ k and F = 21 sin 8. (c) Two-leading shoe :- As for a single-leading shoe, k + l sin8, k-[sin 8 -~ OA = k x r ; AB = 2 k x r ; OP = l x r . A P P E N D I X I D E T E R A M I N A T I O N OF FORMULAE FOR SHOE FACTORS (as defined on pp. 44 and 45) ( 1 ) Factors for sliding shoes with parallel abutments. ON In Fig. 5, drum drag = R X -. r Shoe-tip load = L for leading or M for trailing shoe. R x O N o r R x O N Therefore F = - -L x r M x r - A B x O N o p A B x O N B N x r A N x r 2kl sine 2kl sin ek-1sinOor k + l s i n f l -- (2) Factors for brakes using sliding shoes with parallel abutments. The brake factor is the mean of the shoe factors. The brake layouts are shown in Fig. 9. (a) Floating expander :- 21 cosec 8 cosec2 e - 121k2\u2019 - (b) Fixed cam :- Let x = leading-shoe-tip load, and (2-x) = trailing-shoe-tip load; then (2-x) = leading-shoe-abutment load, and x = trailing-shoe-abutment load"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001770_iros.2015.7353824-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001770_iros.2015.7353824-Figure1-1.png",
        "caption": "Fig. 1. Plan view of two encapsulated ferromagnetic robots R1 and R2, immersed into a water tank.",
        "texts": [
            ", the fluid drag forces opposing robot motion and ~Fi, i.e., the dipole interaction forces neglected in prior work [10], [5]. The force equations are given by: M1~a1 = ~Fd1 + ~Fm1 + ~Fi12 (2a) M2~a2 = ~Fd2 + ~Fm2 + ~Fi21 (2b) where M1, M2 and ~a1, ~a2 are the mass and acceleration terms for each robot R1 and R2, respectively. The robots also experience torques due to the magnetic fields of the MRI and their own interacting fields. Consider that the robot\u2019s magnetic moment ~m forms an angle \u03c6 with respect to the z-axis, as it shown in Fig.1. The following dynamic equations describe the resulting motion: J1\u03c6\u03081 = ~Td1 + ~Tm1 + ~Ti12 (3a) J2\u03c6\u03082 = ~Td2 + ~Tm2 + ~Ti21 (3b) Here J1, J2 are the moments of inertia of the two robots and \u03c61, \u03c62 are the angles that correspond to the magnetic moments ~m1 and ~m2 of robots R1 and R2 with respect to the z-axis, respectively. The terms ~Td1 and ~Td2 represent fluid drag torques. Moreover, ~Tm1 , ~Tm2 are the torques acting on the robots due to the presence of the very strong and uniform magnetic field ~B0 of the MRI scanner, which is directed along the z-axis",
            " To investigate this possibility, it is assumed here that the center of each robot is constrained in the plane, but that each is free to rotate about its center. Under this assumption, (4)-(5) can be used to compute an analytical expression for the equilibrium angles \u03c6?1 and \u03c6?2 of (3a)-(3b), for the two robots R1 and R2, respectively. The tangent of the angle \u03c6?1 is given by: tan\u03c6?1 = 3|| ~m1||\u00b50 sin \u03b8 cos \u03b8 4\u03c0r3||~B0||+ \u00b50|| ~m1||+ 3|| ~m1||\u00b50 sin 2 \u03b8 (6) where ~r is the vector that connects the centers of the robots and forms an angle \u03b8 with respect to the z-axis, as depicted in Fig.1. Using data from Table I of the validation section, we can derive the following equilibrium angles: Considering that ||~B0|| = 3 Tesla, a (minimum) separation distance of ||~r|| = 3.2 mm, and for \u03b8 \u2208 [0, 2\u03c0), it can be found that for a robot of diameter d1 = 1.5mm, the equilibrium angle is \u03c6?1 \u2208 {\u22120.2\u25e6, . . . , 0.2\u25e6}. Since the effect of these small angles on the direction of the magnetic gradient and interaction forces is negligible, robot rotational dynamics will not be considered in the remainder of the paper and, in the force analysis, it will be assumed that the robot dipoles remain aligned with the central field axis, z"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure3-1.png",
        "caption": "Figure 3. Case where batteries in the robot are dead.",
        "texts": [
            " By the above characteristics (i), (iii) and (iv), we can expect that the safety device prevents the robot from colliding with humans at unexpected high speeds (Fig. 1). Furthermore, by (ii), we can adjust the detection velocity level according to the requirement of each patient\u2019s gait exercise. Moreover, by (v), if a human is pressed against a wall by the robot locked by the safety device, we can easily rescue the human by moving the robot in a direction opposite to the direction in which the human is pressed (Fig. 2). Additionally, by (vi), even if the batteries in the robot are dead, the safety device can act because it requires no power supply (Fig. 3). As shown in the next section, each shaft can be locked in clockwise and counterclockwise directions by using two safety devices for each shaft. Also, the characteristics of the torque-based mechanical safety device are as follows: (vii) If the magnitude of the torque of a shaft exceeds a preset threshold level, then the safety device for the shaft switches off all motors of the robot. We call the preset threshold level the \u201cdetection torque level\u201d. (viii) The detection torque level is adjustable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001811_0954405415608784-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001811_0954405415608784-Figure2-1.png",
        "caption": "Figure 2. Test rig layout.",
        "texts": [
            " The hydraulic pressure induces displacement in the axial direction, pushes the outer ring of the rear bearing, and eventually transfers this displacement to the inner ring of the rear bearings through the shaft. Therefore, the displacement caused by hydraulic pressure is converted into axial force. The lubrication system generates an oil\u2013air mixture. The outlets for the oil\u2013air mixture system are connected to three separate channels: front bearings, rear bearings, and load bearings. Throughout the entire process, the amount of oil\u2013air mixture remains constant, which keeps the lubrication system steady. Figure 2 shows the ball bearing instrumentation for the test rig. The highest speed achieved by the motorised spindle is 20,000 r/min. The speed of the motorised high-speed spindle is controlled by a frequency converter. Subsequently, the axial preload and radial loads are adjusted by a proportional hydraulic system. Data on the speed, preload, current, vibration, and temperature are automatically sent to a centralised controller. Vibration in the ball bearing outer ring is measured by an acceleration sensor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001996_sii.2015.7404997-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001996_sii.2015.7404997-Figure7-1.png",
        "caption": "Fig. 7. Initial angle is set to q = [\u2212\u03c0/12, \u03c0/6]T as (a), and a input torque compensating gravity and normal force in such a way as to converge at q = [\u2212\u03c0/3, 2\u03c0/3]T as (b) is set to fi\u0304 = [0,\u2212lmg/4]T.",
        "texts": [
            " Also, because the exerting force described by the coordinate system \u03a30 can be calculated by 0fi = 0Ri ifi, we will confirm whether the acting force 0fi during the restraint motion is correct, and justify the time profile of the acting force and constraint force in this section. First, we examined the acting force and constraint force acting between each link of 2-link manipulator. Constraint condition is z = 0, and is not z \u2264 0. This means the space under z = 0 is invasive, then the shape of 2-link manipulator under z = 0 line is possible in the following simulations. In order to avoid that the equilibrium point of the initial configuration and target one becomes singular, we give Fig.7(a) as the initial shape, also we give \u03c4\u0304 as is given by Eq.(37) as an input to compensate for the gravity and constraint force that can make the position of Fig.7(b) becomes a stable equilibrium point. \u03c4\u0304 = [0,\u2212mgl/4]T (37) Here we set l1 = l2 = l, coefficient of friction is K = 0.01, the initial position as shown in Fig.7(a) is q = [\u22127\u03c0/12, \u03c0/6]. The simulation will not stop until the manipulator is in stationary state(it is determined as to be stationary when the angular velocity of each joint satisfies |q\u0307i| < 0.001 [rad/s]). Fig.8 shows the screen shot of the simulation, Fig.9 shows the time response of the ycomponent of the force 0f1 = [0f1x, 0f1y, 0f1z]T, 0f2 = [0f2x, 0f2y, 0f2z]T acting on each joint and the frictional force ft acting on tip link. Figure 10 shows the time profile of the z-component of the force 0f1, 0f2 and constraint force fn, and Table 1 shows the physical parameters, and Table 2 shows the initial value and converged value for equilibrium configuration"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000708_1.4929331-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000708_1.4929331-Figure10-1.png",
        "caption": "FIG. 10. (a) Calculated swimming trajectories of magnetotactic bacteria in the rotating magnetic field generated by the device described in Figure 5 at \u03b1= 0.115, a= 0.001, \u03d6= 0.2 for \u03c4 < 100; (b) the angular velocity of the bacteria against \u03c4, with T < \u03c4 < 3 T and along the trajectory plotted in (a).",
        "texts": [
            " The magnetic field intensity (normalized with respect to the average value) that satisfies this condition is plotted as a function of \u03b2 (rotation angle of the disc) and is shown in Fig. 9(a). The variations of the calculated normalized field intensity are small. Fig. 9(b) shows how the intensity is affected by a change in \u03b1 by 10% and 20%. A variation of 10% in \u03b1 causes a change in intensity of about 5%. Examples of trajectories calculated according to (7) and (8) for the optimized \u03b1 value are shown in Fig. 10(a). The calculated trajectories are almost circles while in the field of a centrally rotating magnet they were more elliptical (see Fig. 4(a)). The calculated trajectories do not show considerable drift. The small deviations from the circular shape are conditioned by the temporal variation of the angular velocity of the magnetic moment of the bacteria. Fig. 10(b) shows the angular velocity of the bacterium along the trajectory shown in Figure 10(a). Compared with the centrally rotating magnet (see Fig. 4(b)), the width between subsequent peaks and valleys is almost the same and thus produces more circular trajectories. The mechanism that was used to generate the required movements is shown in Figure 11. A Permalloy coin magnet (diameter 3 cm and thickness 1 cm, company: Supermagnete) was used for this experiment. The magnetic field sensor is a GM08 Hirst Gaussmeter with an accuracy of \u00b10.1 Oe. The measurements were performed in a magnetically shielded room"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002986_s00170-016-9507-2-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002986_s00170-016-9507-2-Figure4-1.png",
        "caption": "Fig. 4 Bolted joint elements: a \u201clinear\u201d connection and b \u201carray\u201d connection",
        "texts": [
            " For the \u201clinear\u201d connection form, it is assumed that the dynamic characteristics of the joints between two adjacent bolts are only affected by the mechanical attributes of the two adjacent bolts and have nothing to do with the other bolts, and for the \u201carray\u201d connection, the dynamic characteristics of the joints between four adjacent bolts are only affected by the mechanical attributes of the four adjacent bolts and have nothing to do with the other bolts. The assumptions made above have been proven in a previous study [29]. According to the above assumptions, the joint between each two adjacent bolts is regarded as a bolted joint element in the \u2018linear\u2019 connection form, as shown in Fig. 4a. For the \u201carray\u201d connection, the joint between each four adjacent bolts is regarded as a bolted joint element, as shown in Fig. 4b. Each element has 8 nodes, and every node has 3 translational degrees of freedom (DOFs), so every bolted joint element has 24DOFs. The movements of the joints are represented by relative movements between nodes 1 and 5, 2 and 6, 3 and 7, and 4 and 8. Once the exact relation between the forces on these nodes and their displacements is determined, the dynamic model of the bolted joint is established. n this study, only the elasticity and damping characteristics of the bolted joint are taken into consideration, regardless of the mass"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002232_icma.2014.6885882-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002232_icma.2014.6885882-Figure1-1.png",
        "caption": "Fig. 1. The CAD model of a combined MIS robot.",
        "texts": [
            " In this paper, we try to propose a new kind of preoperative positioning method for a combined surgical robot which is consisted of three same robotic arms, the mid arm is installed with a laparoscope and the other two sides are installed with micro instrument devices; the two sets of solutions of the instrument arms\u2019 preoperative positioning angles are calculated by using numerical computation method, we use the projection length & area determination conditions and the percentage of collaboration workspace to achieve the best preoperative positioning angles of the two instrument arms. This paper focuses on the instrument arms\u2019 preoperative positioning problem of the MISR as shown in Fig. 1, the slave arms of MISR are made up of two instrument arms and one laparoscope arm. 1269978-1-4799-3979-4/14/$31.00 \u00a92014 IEEE The three robotic arms are the same 8 DoFs general robotic arm except for the devices on the end, as shown in Fig. 2, the first four joints are passive and the others are active. Before the operation, the surgical assistants will place the three robotic arms to the appropriate position and orientation, therefore the trocars\u2019 position are fixed to the desired position that is determined by the actual operation requirements"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003407_smc.2016.7844632-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003407_smc.2016.7844632-Figure1-1.png",
        "caption": "Fig. 1. Modified hex rotor platform",
        "texts": [
            " Here, the backstepping approach and thus the non-linear model based attitude control will be implemented on a hex rotor. The model parameters will be identified and a velocity and position control will be added on top of the attitude control applying a cascaded control structure to hex rotors and increasing the autonomy of the vehicle even further. The overall system and the controllers in particular will be validated during real flight experiments to prove the concept and the implementation. The hex rotor platform depicted in Figure 1 is built up of a hex rotor frame. On each cantilever a MK3638 brushless outrunner rotor is mounted, which, together with the 13\u201d propellers, creates the thrust propelling the vehicle. To commute the three phases of the rotors a brushless direct current controller (BLDC) is employed for each rotor. All the former elements are manufactured by the company HiSystems [5]. The flight control unit (FCU) consists of two Beaglebone Black running QNX [6] as real-time operating system. They communicate with each other via serial communication to exchange information needed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000900_s10846-014-0157-z-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000900_s10846-014-0157-z-Figure1-1.png",
        "caption": "Fig. 1 Quad-rotor with its generalized coordinates",
        "texts": [
            " In Section 6 the repulsive scheme is described, while optimization of the controller constants is shown in Section 7. Section 8 presents simulation results and conclusions are given in Section 9. The generalized coordinates for the quad-rotor are the following q = (x, y, z, \u03c6, \u03b8, \u03c8) \u2208 R 6 (1) where \u03be = (x, y, z) \u2208 R 3 denotes the position of the center of mass of the rotorcraft, relative to the inertial frame I, and \u03b7 = (\u03c6, \u03b8, \u03c8) \u2208 R 3 represents the Euler angles (roll, pitch and yaw respectively) that describe the rotorcraft orientation (Fig. 1). The dynamic model of the quad-rotor used in this paper is based on the Euler-Lagrange approach [16] mx\u0308 = u(cos \u03c8 sin \u03b8 cos \u03c6 + sin \u03c8 sin \u03c6) (2) my\u0308 = u(sin \u03c8 sin \u03b8 cos \u03c6 \u2212 cos \u03c8 sin \u03c6) (3) mz\u0308 = u cos \u03b8 cos \u03c6 \u2212 mg (4) \u03c8\u0308 = \u03c4\u03c8 (5) \u03b8\u0308 = \u03c4\u03b8 (6) \u03c6\u0308 = \u03c4\u03c6 (7) 2.1 Yaw and Altitude Control It is desired to maintain a constant yaw orientation \u03c8i = 0 \u2200t > 0, and therefore, the yaw control used for this purpose is \u03c4\u03c8i = \u2212kp\u03c8\u03c8i \u2212 kv\u03c8\u03c8\u0307i (8) where kp\u03c8 and kv\u03c8 denote the proportional and derivative gains of the PD yaw control, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000913_j.mechmachtheory.2014.07.010-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000913_j.mechmachtheory.2014.07.010-Figure9-1.png",
        "caption": "Fig. 9. Deformations of the main shaft in a vertical plane.",
        "texts": [
            " Dynamic deformations in a vertical plane Fig. 8 shows the deformation of the main shaft in a vertical plane O3yz . We can see the deformations of the shaft in three different sections: z = 0[m], z = 0.75[m], and z = 1.5[m]. These graphics are functions of the time t and they change by the harmonic law. We can also determine the amplitudes of the vibrations for the three cross-sections. The greatest deviation is in the initial section, where z = 0. This section is the most dangerous cross-section. Fig. 9 shows the deformations of themain shaft of the band sawmachines. This surface is a function of two arguments: the time t, which changes from 0 to 0.8 s, and coordinate z, which maximum value is equal to the length of the shaft, i.e. zmax = 1.5 [m]. We can drawparallel vertical planes for each value of the coordinate z and see how corresponding cross-sections vibrate, aswell as to measure the amplitudes of the vibrations for these sections. All cross-sections vibrate by the harmonic law. The figure shows at what value of the coordinate z deformation of the shaft are zero"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002999_ictck.2015.7582658-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002999_ictck.2015.7582658-Figure1-1.png",
        "caption": "Figure 1. Model of Two-Wheeled Inverted Pendulum [1]",
        "texts": [
            " Section II presents a dynamic model of a TWIP Newton\u2013Euler approach. Section III is devoted to describe LMI-based RMPC algorithm. The simulation results of the RMPC applied on a TWIP are presented in Section V. Furthermore, comparison of the RMPC with a LQR is brought in this part to show the validity of the proposed approach. Section V presents the conclusions from the study. II. DYNAMIC MODEL In this section, mathematical model of a TWIP using Newton\u2013Euler approach is presented. As depicted in Fig.1, the TWIP consists of two independent actuators (brushless motors) which provide the torques to the wheels to stabilize the robot as well as rotate around the z-axis. The table of forces and moments affecting the vehicle is shown in in Table I. The control objective is to make the heading angle , and the tilt angle converge to desired values, d and zero, respectively [2]. The essential idea is to use the pitch angle as a \u201cgas pedal\u201d for the vehicle and use it to accelerate and decelerate until the specified speed is attained [15]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003074_icarm.2016.7606961-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003074_icarm.2016.7606961-Figure1-1.png",
        "caption": "Fig. 1. A three-dimensional FWMAV model",
        "texts": [
            " 2Also the Fundamental Research Funds for the China Central Universities of USTB under Grant FRF-TP-15-005C1. reconnaissances, search and rescue missions in disaster areas or battlefields, even people to play with in the future [4\u20137]. For studying and realizing the trajectory tracking control of FWMAV well, the nonlinear dynamics of a rigid bodyFWMAV should be modeled according to the desires. First of all, multiple frames should be defined. Among them, there are three important reference frames: wing frame, stroke plane and body frame as Fig. 1. For designing the attitude and position controllers, many parameters of the systems are measured and established with respect to them [4]. Secondly, relative orientations should be determined by different rotation matrices. On the basis of inspirations about insects and micro birds flying, each wing of FWMAV is founded by three degrees of freedom. What is more, the major movements of wing are established via stroke plane [8, 9]. The wings are able to twist and translate at the wing root and adjust the angle of pitch of body on the basis of stroke plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000682_c5ra10802k-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000682_c5ra10802k-Figure2-1.png",
        "caption": "Fig. 2 The detection principle of the immunosensor.",
        "texts": [
            "2 mg mL 1 PA (pH 3) aqueous solution for 2 h to combine PA through electrostatic adsorption, then incubated with 100 ng mL 1 Ab1 solution for 12 h at 4 C, and nally rinsed with water and stored at 4 C for use. Each procedure was followed by careful washing. Based on a sandwich-type immunoassay format, the immunosensor was rst incubated with the target antigen for 40 min at 37 C, then incubated with 20 mL of Ab2 composite for 40 min at 37 C, and acted as the working electrode and was detected by DPV scan in 0.1 M PBS (pH 7.0) containing 5 mM AA. The following catalysis reaction between the Ab2 composite and AA would happen (Fig. 2). This journal is \u00a9 The Royal Society of Chemistry 2015 With this reaction, the response current could be greatly amplied, and the current intensity was quantitatively related to the content of target antigen. The CV characterization of the fabrication procedure in 5.0 mM K3[Fe(CN)6]/K4[Fe(CN)6] solution is presented in Fig. 3. Typical reversible redox peaks of ferricyanide ions could be observed for the bare GCE (Fig. 3A, curve a); aer electrodeposition of AuNPs on the surface of the GCE, the redox peaks obviously increased (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001536_exsy.12115-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001536_exsy.12115-Figure1-1.png",
        "caption": "Figure 1: Diagram of the acrobot.",
        "texts": [
            " In Section 2, the details of the acrobot control problem are presented; then we describe the GP approach taken here and how it was applied to the swing up and balance problems in Section 3; this is followed by the results of the two experiments (Section 4); and finally, in Section 5, we present our conclusions and directions of future research. \u00a9 2015 Wiley Publishing Ltd Expert Systems, xxxx 2015, Vol. 00, No. 00 The acrobot is a two-link robot, based on a human acrobat. The links are connected by an actuated joint, and one end is connected to a bar by an unactuated joint (Spong, 1994) (Figure 1). The state x \u00bc \u03b81; \u03b82; _\u03b81; _\u03b82 \u22a4 comprises the first angle \u03b81, second angle \u03b82 and the two angular velocities _\u03b81 and _\u03b82 . The control variable is the torque \u03c4 to be applied to the actuated joint. The acrobot is a difficult control task as it is a fourdimensional, highly non-linear, under-actuated control problem (Spong, 1994, 1995; RLC, 2009). The task of the acrobot is to swing up, in the minimum time, from the initial state \u2192x0 \u00bc 0; 0; 0; 0\u00bd \u22a4 to the inverted, unstable, target state \u2192xt \u00bc \u03c0; 0; 0; 0\u00bd \u22a4 and then to remain balanced in that position"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.43-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.43-1.png",
        "caption": "FIGURE 8.43",
        "texts": [
            "1, are included in the tutorial lessons. The sliding block example is simply prepared for an easy start with both SolidWorks Motion and Pro/ENGINEERMechanism Design. Default options and values are mostly used. Once readers are more familiar with either of these two software tools, they may move to the second tutorial example, the single-piston engine. The first example simulates a block sliding down a 30-deg. slope with no friction. Because of gravity, the block slides and hits the ground, as depicted in Figure 8.43. Simulation results obtained from motion software can be verified using particle dynamics theory learned in a physics class. Object path on a large-scale waterslide showing critical areas of safety concern. The physical model of a sliding block is very simple. The block was made of cast alloy steel with a size of 10 in. 10 in. 10 in. As shown in Figure 8.43, it traveled a total of 9 in. The units system employed for this example was IPS (inch, pound, second). The gravitational acceleration was 386 in./sec2. The block was released from a rest position (that is, the initial velocity is zero). The block and slope (or ground) were assumed rigid. A limit distance mate (in SolidWorks) was defined to prevent the block from sliding out of the slope face. The block reached the end of the slope face in about 0.3 sec, as indicated in Figure 8.44(a), which shows the Y-position of the mass center of Sliding block: (a) schematic view and (b) motion model in CAD (SolidWorks)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000991_scis-isis.2014.7044740-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000991_scis-isis.2014.7044740-Figure1-1.png",
        "caption": "Fig. 1. Porter Robot based on Wheeled Inverted Pendulum System",
        "texts": [
            " The inverted pendulum mobile robot has many advantages over statically stable wheeled robots and other biped robots. It requires less space to stand and stay upright than ordinary wheeled mobile robots and a smaller number of actuators than conventional biped robots. They support human locomotion and/or small goods transportation based on inverted pendulum upright controllers[4][5]. We have also developed a suitcase-type two-wheeled inverted pendulum mobile robot that changes the control parameters according to the user intention[1]. Fig.1 shows a concept of a porter robot based on a wheeled inverted pendulum mobile system. The robot stands by itself if the user leaves it alone. The user operates the robot by pulling the handle, then, the robot assists the power of the pulling and follows the user. It is useful especially for elderly people or disease patients because it needs less power to transport a heavy equipment for the user. The most conventional inverted pendulum mobile robot controls to follow the fixed desired posture angle and wheel velocity",
            " The selector has to change the desired posture and desired wheel velocity according to the user intention/behavior appropriately. When the user puts the robot in upright position, the desired posture should be at the right balance point otherwise the robot starts to run. The desired wheel velocity and the control gain corresponding to the velocity are set to zero and high value, respectively, in order to keep the robot to stay. When the user starts to pull the robot, the desired position changes to incline the robot body to the user for comfortable operation of the robot as shown in Fig.1. The control gain corresponding to the wheel velocity is set to zero so that that robot follows the user. III. INTENTION RECOGNITION MODULE An intention recognition module has a motion template that corresponds to one intention of the user. The motion template maintains a sequence of sensory outputs and motor inputs \u03bct and their variances \u03c3t that corresponds to the assigned intention. The center of the template \u03bct and the variance \u03c3t include the estimated body posture angle based on the outputs of two-directional accelerometer xa, one rate gyroscope xg, wheel encoder output xe, and input to the wheel motor xu at time t"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002919_1.4962785-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002919_1.4962785-Figure6-1.png",
        "caption": "Fig. 6. Schematic of the angles and parameters as the ball hits the inside of the right goalpost.",
        "texts": [
            " 54, OctOber 2016 435 For other values of x, the ball moves toward the right side of the soccer field, following the goal line at b = 0\u02da. For all negative values of b the ball moves toward the back of the net. From Fig. 4, Eq. (3), and a trigonometric identity, we find that the y-location of the ball on the left post can be calculated as (7) \u2022 Double post bounce? Next, we will analyze the motion of the ball after it bounces off the left post and moves toward the right post. The ball travels the length of the goal, reaching the right post at y2 from its center (see Fig. 6). Using Eqs. (6) and (7), we get (8) The condition for a double post bounce is that \u2013(RP + RB ) < y2 < (RP + RB). (9) Figure 7, obtained from Eq. (8), shows that this is satisfied for a range Dxd of approximately 0.6 mm, which according to Eq. (6) corresponds to a Db of 2.5\u02da. Using Eq. (3) we can calculate the corresponding q for each x and divide the inside quarter of the left goalpost into regions of different ball impact locations (see Fig. 8). The small margin for a double post bounce shows why one does not see it very frequently in soccer. The Dq for the double post bounce for a penalty is indicated in green in Fig. 8 and can be calculated from Db = 2.5\u02da and Eq. (3) to be only about 1.2\u02da. \u2022 Ricochet off the right post The angle to the goal line, e, relates to b and j as (see Fig. 6) e = 2j \u2013 b . (10) Further, the angles j, b, and s relate as j = b + s. (11) The angle s can be calculated from y2 and the dimensions of the ball and goalpost as (12) Inserting Eqs. (11) and (12) into Eq. (10), we get (13) Using Eq. (13), Fig. 9 shows the angle e for different values of x. For e > 90\u02da the ball bounces off to the right side of Inserting Eqs. (2), (3), and (4) into Eq. (5), we get b as a function of the distance from the contact point to the center of the post x as (6) Note that since a, b >> x, we ignore the slight x dependence of a in Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.10-1.png",
        "caption": "Fig. 3.10 Examples for complex joints with physical contact surfaces",
        "texts": [
            " As a consequence, the spherical motion is comprised of three revolute joints with intersecting axes in the fixed point O of the kinematic chain (Fig. 3.9). In mechanisms and gear trains there is a difference between standard joints (socalled lower kinematic pairs) and complex joints (higher kinematic pairs) (Reuleaux 1875): \u2022 In standard joints, the bodies have surface contact. One distinguishes the following six standard joints (Table 3.3). \u2022 Complex joints have contact along a body line or at a point (Fig. 3.10). Contact between two bodies can occur on two non-physical, spatially fixed surfaces or rather body fixed surface, spur surface or rather pin surface, along the instantaneous axis of rotation or rather screw axis, which indicate the instantaneous motion state. In the suspending motion of a five-link wheel suspension, for example, the motion of the wheel carrier relative to the chassis can be represented as a screw motion of the wheel carrier-fixed pin surface with respect to the chassis-fixed spur surface"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003415_ecce.2016.7854897-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003415_ecce.2016.7854897-Figure5-1.png",
        "caption": "Fig. 5. Schematic diagram of current signal acquisition from a Skystream 3.7 wind turbine.",
        "texts": [
            " The variation of fc(t) can be easily estimated from the time-frequency distribution (TFD) of the signal using the ridge searching algorithm. Then the SRF is calculated by ( ) ( ) /r cf t f t q= (25) where q is the number of pole pairs of the PMSG. The details of the instantaneous SRF estimation from the PMSG stator current signal can be found in [5]. The generator current signal of a 2.4 kW Skystream 3.7 wind turbine is first analyzed by the proposed MFS method for bearing fault diagnosis. The schematic diagram of the current signal acquisition is illustrated in Fig. 5. It is a wireless sensor network-based data acquisition system. One phase current signal is measured by a current sensor at the stator terminal of the PMSG of the wind turbine and is sampled by a wireless sensor node (Model: V-Link\u00ae -LXRS\u00ae) at 1000 Hz. The sensor node is installed in a box attached to the bottom of the wind turbine. The current data is then sent wirelessly to a gateway (Model: WSDA\u00ae -1500 -LXRS\u00ae) located within two kilometers, which sends the data through Ethernet to the SensorCloud, which is a network server for data storage"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002383_s13534-015-0188-9-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002383_s13534-015-0188-9-Figure1-1.png",
        "caption": "Fig. 1. Design and configuration of RehabWheel developed in curent study.",
        "texts": [
            " Therefore, in this study, we aimed to suggest a new concept of a mobility system that integrates a wheelchair platform, used for the implementation of mobility, and a wearable exoskeleton structure, used for the implementation of gait assistance, and to validate its structural stability and gait control confidence. Design of system structure and analysis of structural stability The entire system presented in this research was designed for the purpose of developing a novel mobility device, which supports mobility in Activities of daily living (ADLs), and function of gait assistance for patients with gait disorders (Fig. 1). The developed system was named the RehabWheel. The basic structure of the RehabWheel was designed based on the platform of a wheelchair for the implementation of mobility, and incorporated a wearable exoskeleton structure formed with six one-degree-of-freedom (DoF) rotational joints for implementing the function of gait assistance (Fig. 1). In other words, it was designed to facilitate the functions of moving and walking together in a single system. In the case of the wheelchair platform, the design alows for both the mobility mode, which provides movement as a means of transportation while siting, and the walk assistance mode, which provides gait assistance by synchronizing with the moving speed of the wheels on the wheelchair platform when standing. In the case of the wearable exoskeleton part, 12 pneumatic artificial muscles (30 mm Air Muscle, Shadow Robot Company, UK) were atached in the design for the 1- DoF flexion-extension operation of the hip, knee, and ankle joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003267_physrevc.94.065805-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003267_physrevc.94.065805-Figure4-1.png",
        "caption": "FIG. 4. (a) The axially symmetric deformation of cylindrical cluster. (b) shows the charge density profile for a fixed \u03c6,z.",
        "texts": [
            " (A1) The left-hand side (LHS) of the above equation is 0 because the contribution from the normal derivative of or \u03b4 taken for opposite sides of the cell are canceled, for example for the side perpendicular to the x axis \u2202n |x=\u2212a/2 = \u2212\u2202n |x=a/2 (A2) and similarly for the other cell sides. In the case of an isolated cell (sphere, cylinder) instead of periodic boundary conditions the potential and its variation obeys the Neumann boundary condition \u2202n |\u2202C = 0 (A3) and again the LHS of Eq. (A1) vanishes. SURFACE INTEGRALS Here we illustrate how the charge variation \u03b4\u03c1 caused by the change in cluster shape \u03b5 may be expressed by the surface integral on an unperturbed cluster surface. As an example let us consider a cylindrical cluster with radius R perturbed by a deformation \u03b5, see Fig. 4. The deformation \u03b5(z,\u03c6) is a function defined on the cylinder surface r =R. In cylindrical coordinates r,\u03c6,z the charge distribution, before and after deformation, is given by \u03c1(r,\u03c6,z) = \u03c1\u2212 + (\u03c1+ \u2212 \u03c1\u2212) \u03b8 (R \u2212 r), (B1) \u03c1 \u2032(r,\u03c6,z) = \u03c1\u2212 + (\u03c1+ \u2212 \u03c1\u2212) \u03b8 (R + \u03b5 \u2212 r). (B2) By use of the Taylor expansion in powers of \u03b5(z,\u03c6) one gets \u03b4\u03c1(r,\u03c6,z) = \u03c1 \u03b4(R \u2212 r) \u03b5(z,\u03c6) \u2212 1 2 \u03c1 \u03b4\u2032(R \u2212 r) \u03b5(z,\u03c6)2. (B3) The expression for electrostatic energy or perturbed potential is always represented by a linear functional of \u03b4\u03c1 integrated with some function f [Eqs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003294_vppc.2016.7791805-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003294_vppc.2016.7791805-Figure1-1.png",
        "caption": "Fig. 1. Five phase 10-slot/8-pole SPM machines with different teeth structure",
        "texts": [],
        "surrounding_texts": [
            "The 10-slot/8-pole five phase SPM machines with equal and unequal tooth width are employed to investigate the influence of the stator structures on the output torque considering the third harmonic current injection. Figs.1 (a) and (b) show the stator and winding connections. The stator parameters of both equal and unequal tooth machines have been optimized to maximize the output torque with the constraint of the same copper loss and rotor diameter considering the third harmonic current injection. All the stator, PM, and rotor parameters are shown in TABLE I. FE VALIDATION In five phase machine, the ideal sinusoidal and sinusoidal with third harmonic currents are the functions of rotor position \u03b8, as expressed as: ( ) sin( ) ( ) [sin( ) sin(3 )] m a i I p i I p a p \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 = = + (1) where p is the number of pole-pair, and Ia and a are the parameters need to be determined. a is the ratio of the injected third harmonic current to the fundamental one, while Ia varies with the injected third harmonic in order to maintain the same rms as the sinusoidal phase current Im. In order to have the same rms value as the sinusoidal current, Ia and Im should be: 2 2 2 2 ma aI a I I+ = (2) For five phase machine, the average output torque considering the third harmonic current can be expressed as: ( )1 1 3 35 = E I E I T + \u03a9 (3) where 1E and 3E are the fundamental and the third harmonic phase back-EMFs. 1I and 3I are the fundamental and the third harmonic phase currents. \u03a9 is the angular speed. Thus the average output torque with and without the third harmonic current can be given as: 1 1 1 3 2 2 2 5 5 5 1 1= m m m E I T E aE I I a aT = \u03a9 + + + \u03a9 (4) In order to obtain the torque improvement, 1T and 2T should be: 2 2 1 3 1 2 2 5 1 1 0 1 1 mI a a T T E E a a  + \u2212  \u2212 = \u2212 >  \u03a9 + +  (5) The value of a can be obtain as: 1 3 2 2 1 3 2E E a E E < \u2212 (6) It can be seen that the output torque can be improved with the third harmonic current injection when a satisfies (6). Otherwise, the torque would not be increased even under the condition of the third harmonic current injection and this ratio of the third harmonic current to the fundamental one is defined as critical ratio. The maximum output torque can be determined by differentiating 2T with respect to a and equating it to zero, as expressed as: ( ) ( ) 1 2 2 23 1 32 2 1 1 0 1 E a a a E aEdT da a \u2212 + \u2212 + + = = + (7) From (4) and (7), the values of a and 2T can be obtained as: 3 1 E a E = (8) 2 2 1 3 2 2 2 2 1 3 1 3 2 2 2 1 3 5 5 = 5 m m m E E I I E E E ET E E I + + + \u03a9 += \u03a9 (9) Thus, from (4) and (9), the torque improvement can be obtained as: 2 2 1 3 2 1 1 1 1 T T E T E \u2212 = + \u2212 (10) Therefore, the optimal ratio and output torque can be evaluated by (8) and (9), as can be seen that the torque improvement is only dependent on the ratio of the third harmonic back-EMF to the fundamental one. The variation of the torque improvement with the ratio of the third harmonic back-EMF to the fundamental one is shown in Fig. 2. The optimal ratio and torque improvement increase with the increase of the third harmonic back-EMF. However, the torque improvement is not remarkable when the ratio is less than 0.2. In order to validate the critical and optimal injected value of third harmonic current, the influence of the ratio of the third harmonic current to the fundamental one on the output torque is investigated by FE analyses. Fig. 3 shows the variation of the average output torque for the 10-slot/8-pole equal tooth and unequal tooth SPM machines with the ratio of the third harmonic current to the fundamental one. It can be seen that the unequal tooth machine has higher optimal injected third harmonic current and torque improvement than the equal tooth machine since the unequal tooth machine has higher third harmonic back-EMF, which will be discussed in the following section. IV. ELECTROMAGNETIC PERFORMANCE OF EQUAL AND"
        ]
    },
    {
        "image_filename": "designv11_64_0002380_transjsme.15-00563-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002380_transjsme.15-00563-Figure4-1.png",
        "caption": "Fig. 4 Kinematics around the foldable arm",
        "texts": [],
        "surrounding_texts": [
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u306f\u3053\u308c\u306b\u52a0\u3048\uff0c\u8155\u90e8\u306e\u6298\u308a\u305f\u305f\u307f\u65b9\u5411\u3068\u30d2\u30f3\u30b8\u69cb\u9020\u306b\u7740\u76ee\u3059\u308b\u3053\u3068\u3067\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306b\u3088\u308b\u81ea\u5df1\u5c55\u958b\u304c\u53ef\u80fd\u306a\u69cb \u9020\u3092\u63d0\u6848\u3059\u308b\uff0e\u63d0\u6848\u3059\u308b\u69cb\u9020\u3067\u306f\uff0c\u5c55\u958b\u30fb\u56fa\u5b9a\u306e\u305f\u3081\u306e\u98db\u884c\u306b\u76f4\u63a5\u95a2\u4fc2\u7121\u3044\u30a2\u30af\u30c1\u30e5\u30a8\u30fc\u30bf\u304c\u4e0d\u8981\u3068\u306a\u308a\uff0c\u6a5f\u4f53\u91cd \u91cf\u3092\u6291\u3048\u3089\u308c\u308b\uff0e\n\u63d0\u6848\u306b\u57fa\u304d\u8a2d\u8a08\u3057\u305f\u30d2\u30f3\u30b8\u90e8\u306e\u8a73\u7d30\u3092\u56f3 3\u306b\u793a\u3059\uff0e\u3053\u306e\u30d2\u30f3\u30b8\u306f\u30b9\u30e0\u30fc\u30ba\u306b\u53d7\u52d5\u52d5\u4f5c\u3059\u308b\u305f\u3081\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u304c \u52d5\u4f5c\u3057\u3066\u3044\u306a\u3044\u5834\u5408\uff0c\u91cd\u529b\u306b\u3088\u308a\uff0c\u8155\u90e8\u306f\u4e0b\u306b\u6298\u308c\u66f2\u304c\u308b\uff08\u56f3 3-\u5de6\uff09\uff0e\u4e00\u65b9\uff0c\u30d7\u30ed\u30da\u30e9\u306e\u63a8\u529b\u65b9\u5411\u3068\uff0c\u8155\u90e8\u306e\u5c55\u958b \u65b9\u5411\u3092\u4e00\u81f4\u3055\u305b\u308b\u3053\u3068\u306b\u3088\u308a\uff0c\u98db\u884c\u7528\u30e2\u30fc\u30bf\u3092\u56de\u8ee2\u3055\u305b\u3066\u63a8\u529b\u3092\u5f97\u308b\u3053\u3068\u3067\uff0c\u8155\u90e8\u304c\u81ea\u7136\u306b\u5c55\u958b\u3059\u308b\uff08\u56f3 3-\u4e2d\uff09\uff0e \u8155\u90e8\u306f\uff0c\u6a5f\u4f53\u30d9\u30fc\u30b9\u90e8\u3068\u540c\u4e00\u5e73\u9762\u4e0a\u3068\u306a\u308b\u5730\u70b9\u3067\uff0c\u30b9\u30c8\u30c3\u30d1\u306b\u3088\u308a\u5c55\u958b\u304c\u5236\u9650\u3055\u308c\uff0c\u4e00\u822c\u7684\u306a\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u5f62\u72b6 \u3068\u306a\u308b\uff08\u56f3 3-\u53f3\uff09\uff0e\u8155\u90e8\u306e\u7a81\u8d77\u3068\u30b9\u30c8\u30c3\u30d1\u304c\u5d4c\u307e\u308a\u5408\u3046\u3053\u3068\u3067\uff0c\u30d2\u30f3\u30b8\u90e8\u306e\u904a\u3073\u306b\u3088\u308b\u8155\u90e8\u306e\u306d\u3058\u308c\u30fb\u632f\u52d5\u3092\u9632\u3050\uff0e\n\u306a\u304a\uff0c\u672c\u8ad6\u6587\u3067\u306f\u6298\u308a\u305f\u305f\u307f\u6a5f\u69cb\u306e\u6709\u7528\u6027\u3092\u691c\u8a3c\u3059\u308b\u3053\u3068\u3092\u76ee\u7684\u3068\u3059\u308b\u305f\u3081\uff0c\u6a5f\u4f53\u306b\u306f\u5b9f\u969b\u306e\u63a2\u67fb\u306b\u5229\u7528\u3059\u308b\n\u30bb\u30f3\u30b5\u3084\uff0c\u5916\u58c1\u3068\u306e\u63a5\u89e6\u9632\u6b62\u7528\u306e\u30d7\u30ed\u30da\u30e9\u30ac\u30fc\u30c9\u7b49\u306f\u642d\u8f09\u3057\u306a\u3044\uff0e\n2\u00b72\u00b71 \u8155\u90e8\u306e\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u306e\u691c\u8a3c\n\u8155\u90e8\u5c55\u958b\u306b\u5fc5\u8981\u306a\u63a8\u529b\u3092\u691c\u8a3c\u3059\u308b\uff0e\u56f3 4\u306b\u53ef\u52d5\u8155\u90e8\u306b\u50cd\u304f\u529b\u3092\u793a\u3059\uff0e\u3053\u306e\u56f3\u306b\u304a\u3044\u3066\uff0c\u30d2\u30f3\u30b8\u306e\u56de\u8ee2\u4e2d\u5fc3\u3067\u3042\u308b\n\u70b9 Q\u56de\u308a\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u306e\u5f0f\u306f\u6b21\u5f0f\u306e\u3068\u304a\u308a\u3068\u306a\u308b\uff0e\u305f\u3060\u3057\uff0cl f \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u30d7\u30ed\u30da\u30e9\u4e2d\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0c l f g \u3092\u30d2\u30f3\u30b8\u56de\u8ee2\u4e2d\u5fc3\u304b\u3089\u53ef\u52d5\u8155\u90e8\u91cd\u5fc3\u307e\u3067\u306e\u9577\u3055\uff0cm\u3092\u53ef\u52d5\u8155\u90e8\u306e\u8cea\u91cf\uff0cFm \u3092\u30d7\u30ed\u30da\u30e9\u63a8\u529b\uff0c\u03b8 f \u3092\u53ef\u52d5\u8155\u90e8\u306e \u6298\u308a\u305f\u305f\u307f\u89d2\u5ea6\uff0cg\u3092\u91cd\u529b\u52a0\u901f\u5ea6\u3068\u3059\u308b\uff0e\u306a\u304a\uff0c\u7c21\u5358\u306e\u305f\u3081\uff0c\u30d7\u30ed\u30da\u30e9\u63a8\u529b\u306f\u8155\u90e8\u4e0a\u3067\u529b\u3092\u767a\u63ee\u3059\u308b\u3082\u306e\u3068\u3059\u308b\uff0e\n0 =\u2212l f gmgcos\u03b8 f + l f Fm (1)\n\u3053\u306e\u5f0f\u3092 Fm \u306b\u3064\u3044\u3066\u307e\u3068\u3081\u308b\u3068\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm = l f g\nl f mgcos\u03b8 f (2)",
            "\u00a9 2016 The Japan Society of Mechanical Engineers[DOI: 10.1299/transjsme.15-00563]\n\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3057\u3066\u3044\u308b\u5834\u5408\uff0c\u03b8 f = 0\u3067\u3042\u308b\u305f\u3081\uff0c\u53ef\u52d5\u8155\u90e8\u3092\u5b8c\u5168\u5c55\u958b\u72b6\u614b\u306b\u4fdd\u3064\u305f\u3081\u306e\u6761\u4ef6\u306f\u6b21\u5f0f\u3068\u306a\u308b\uff0e\nFm \u2265 l f g\nl f mg (3)\n\u4e00\u65b9\u3067\uff0c\u6a5f\u4f53\u304c\u30db\u30d0\u30ea\u30f3\u30b0\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u6a5f\u4f53\u306e\u5168\u8cea\u91cf\u3092\u652f\u3048\u308b\u3060\u3051\u306e\u63a8\u529b\u304c\u5fc5\u8981\u3068\u306a\u308b\uff0e\u30db\u30d0\u30ea\u30f3\u30b0\u6642\uff0c4\u3064 \u306e\u30d7\u30ed\u30da\u30e9\u304c\u3059\u3079\u3066\u540c\u3058\u63a8\u529b\u3092\u767a\u63ee\u3057\u3066\u3044\u308b\u3068\u4eee\u5b9a\u3059\u308b\u3068\uff0c\u6a5f\u4f53\u5168\u4f53\u306e\u529b\u306e\u91e3\u308a\u5408\u3044\u304b\u3089\u30d7\u30ed\u30da\u30e9\u3042\u305f\u308a\u306e\u63a8\u529b Fm \u3092\u6c42\u3081\u308b\u3068\u6b21\u5f0f\u3067\u8868\u73fe\u3067\u304d\u308b\uff0e\u305f\u3060\u3057\uff0cmb \u3092\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u8155\u90e8\u4ee5\u5916\u306e\u6a5f\u4f53\u8cea\u91cf (\u30d9\u30fc\u30b9\u90e8\u8cea\u91cf)\u3068\u3059\u308b\uff0e\n4Fm = (mb +4m)g\nFm = ( 1 4 mb +m ) g (4)\n\u3053\u306e\u5f0f\u3068\uff0c\u53ef\u52d5\u8155\u90e8\u304c\u5b8c\u5168\u5c55\u958b\u3059\u308b\u305f\u3081\u306e\u6761\u4ef6\u5f0f (3)\u304b\u3089\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306b\u53ef\u52d5\u8155\u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u306a\u3044\u305f\u3081\u306e\u6761 \u4ef6\u304c\u6b21\u5f0f\u3068\u306a\u308b\uff0e( 1 4 mb +m ) g \u2265 l f g l f mg\nmb \u2265 4m ( l f g\nl f \u22121\n) (5)\n\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3067\u306f\uff0c\u53ef\u52d5\u8155\u90e8\u306e\u5148\u7aef\u306b\u6700\u91cd\u91cf\u7269\u306e\u30e2\u30fc\u30bf\u304c\u3064\u3044\u3066\u304a\u308a\uff0c\u305d\u306e\u70b9\u3067\u63a8\u529b\u304c\u767a\u63ee\u3055\u308c\u308b\uff0e\u305d\u306e \u305f\u3081\uff0cl f g < l f \u3068\u306a\u308a\uff0c\u5f0f (5)\u306e\u53f3\u8fba\u306f\u5e38\u306b\u8ca0\u3068\u306a\u308b\u3053\u3068\u304b\u3089\uff0c\u5e38\u306b\u6210\u308a\u7acb\u3064\uff0e\u3053\u308c\u3088\u308a\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u98db\u884c\u6642\u306b\u8155 \u90e8\u304c\u6298\u308a\u305f\u305f\u307e\u308c\u308b\u3053\u3068\u306f\u7121\u3044\u3053\u3068\u304c\u8a3c\u660e\u3055\u308c\u305f\uff0e\n2\u00b72\u00b72 \u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u6642\u306e\u52d5\u4f5c\n\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u306f\u3059\u3079\u3066\u306e\u30d7\u30ed\u30da\u30e9\u304c\u307b\u307c\u304a\u306a\u3058\u56de\u8ee2\u6570\u30fb\u63a8\u529b\u3092\u767a\u751f\u3057\u3066\u3044\u308b\u304c\uff0c\u79fb\u52d5\u306e\u305f\u3081 \u306b\u306f\u5404\u30d7\u30ed\u30da\u30e9\u3067\u63a8\u529b\u5dee\u3092\u4f5c\u308b\u3053\u3068\u3067\u6a5f\u4f53\u3092\u50be\u3051\uff0c\u6c34\u5e73\u65b9\u5411\u306e\u63a8\u529b\u3092\u5f97\u308b\uff0e\u307e\u305f\uff0c\u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u306f\uff0c\u5404\u30d7\u30ed\u30da\u30e9\u306b \u3088\u308a\u767a\u751f\u3059\u308b\u53cd\u30c8\u30eb\u30af\u3092\u5229\u7528\u3057\u3066\u5b9f\u73fe\u3059\u308b\uff0e\u305d\u306e\u305f\u3081\uff0c\u79fb\u52d5\u6642\u3084\u30e8\u30fc\u56de\u8ee2\u6642\u306b\u306f\uff0c\u30db\u30d0\u30ea\u30f3\u30b0\u6642\u3088\u308a\u3082\u63a8\u529b\u304c\u5c0f \u3055\u304f\u306a\u308b\u30d7\u30ed\u30da\u30e9\u304c\u5b58\u5728\u3059\u308b\u3053\u3068\u306b\u306a\u308a\uff0c\u524d\u8ff0\u306e\u8155\u90e8\u3092\u4fdd\u3064\u6761\u4ef6\u3092\u6e80\u305f\u3059\u3053\u3068\u304c\u3067\u304d\u306a\u3044\u5834\u5408\u304c\u5b58\u5728\u3059\u308b\uff0e\u7279\u306b",
            "\u6a5f\u4f53\u306e\u30e8\u30fc\u56de\u8ee2\u52d5\u4f5c\u306f\uff0c\u63a8\u529b\u306b\u5bfe\u3057\u3066\u5927\u5e45\u306b\u5c0f\u3055\u3044\u53cd\u30c8\u30eb\u30af\u3092\u7528\u3044\u308b\u305f\u3081\uff0c\u5404\u30d7\u30ed\u30da\u30e9\u3067\u306e\u56de\u8ee2\u6570\u5dee\u3092\u5927\u304d\u304f\u3059 \u308b\u5fc5\u8981\u304c\u3042\u308a\uff0c\u7279\u306b\u6ce8\u610f\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\n\u5177\u4f53\u7684\u306b\uff0c\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u6a5f\u4f53\u30e8\u30fc\u56de\u8ee2\u52d5\u4f5c\u306f\uff0c\u56f3 5\u306e\u3088\u3046\u306b\u56de\u8ee2\u65b9\u5411\u306e\u7570\u306a\u308b\u30d7\u30ed\u30da\u30e9\u540c\u58eb\u306e\u56de\u8ee2\u6570\u306b\u5dee\u3092 \u3064\u3051\u308b\u3053\u3068\u3067\u5b9f\u73fe\u3059\u308b (Mahony et al., 2012)\uff0e\u4f8b\u3048\u3070\uff0c\u56f3 5(a)\u306b\u304a\u3044\u3066 CCW\u56de\u8ee2\u3059\u308b\u305f\u3081\u306b\u306f\uff0c\u56f3 5(b)\u306e\u3088\u3046\u306b CW\u56de\u8ee2\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u3092\u4e0a\u3052\uff0cCCW\u56de\u8ee2\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u3092\u4e0b\u3052\u308b\u3053\u3068\u3067\u5b9f\u73fe\u3055\u308c\u308b\uff0e\u3053\u306e\u6642\uff0c\u305d\u308c\u305e\u308c\u306e \u56de\u8ee2\u6570\u304c\u5909\u5316\u3057\u3066\u3082\uff0c\u5408\u8a08\u306e\u63a8\u529b\u304c\u5909\u5316\u3057\u306a\u3044\u3088\u3046\u306b\u5236\u5fa1\u3059\u308b\u3053\u3068\u3067\uff0c\u6a5f\u4f53\u306e\u9ad8\u5ea6\u3092\u4fdd\u3064\uff0e\u6a5f\u4f53\u306e\u5927\u304d\u306a\u30e8\u30fc\u89d2 \u52a0\u901f\u5ea6\u3092\u5f97\u308b\u305f\u3081\u306b\u306f\uff0c\u305d\u308c\u305e\u308c\u306e\u56de\u8ee2\u65b9\u5411\u306e\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u5dee\u3092\u5927\u304d\u304f\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\u307e\u305f\uff0c\u6a5f\u4f53\u306b\u50cd\u304f \u7a7a\u6c17\u62b5\u6297\u306b\u3088\u308a\uff0c\u9ad8\u901f\u30e8\u30fc\u56de\u8ee2\u3059\u308b\u5834\u5408\u306b\u306f\uff0c\u30d7\u30ed\u30da\u30e9\u306e\u56de\u8ee2\u6570\u5dee\u3092\u7dad\u6301\u3059\u308b\u5fc5\u8981\u304c\u3042\u308b\uff0e\u3060\u304c\uff0c\u4e00\u822c\u306b\u30d7\u30ed\u30da\u30e9 \u306e\u56de\u8ee2\u6570\u5dee\u3068\u6a5f\u4f53\u306e\u30e8\u30fc\u56de\u8ee2\u901f\u5ea6\u306e\u95a2\u4fc2\u3092\u7406\u8ad6\u5024\u3088\u308a\u6c42\u3081\u308b\u3053\u3068\u306f\u56f0\u96e3\u3067\u3042\u308b\u305f\u3081\uff0c\u5b9f\u6a5f\u306b\u3088\u308a\u5b9f\u6e2c\u3057\u691c\u8a3c\u3059\u308b \u5fc5\u8981\u304c\u3042\u308b\uff0e\n2\u00b73 \u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u306e\u5b9f\u88c5\n\u63d0\u6848\u624b\u6cd5\u3092\u57fa\u306b\uff0c\u53ef\u6298\u578b\u30de\u30eb\u30c1\u30b3\u30d7\u30bf\u3092\u5b9f\u88c5\u3057\u305f\uff0e\u56f3 6\u306b\u6a5f\u4f53\u306e\u5916\u89b3\u3092\uff0c\u8868 1\u306b\u6a5f\u4f53\u8af8\u5143\u3092\u793a\u3059\uff0e \u6a5f\u4f53\u306e\u6700\u5c0f\u5e45\u306f\uff0c\u5c55\u958b\u6642\u3067 618 mm\u3067\u3042\u308a\uff0c\u6298\u308a\u305f\u305f\u3080\u3068\u305d\u306e 48.5%\u7a0b\u5ea6\u3067\u3042\u308b 300 mm\u3068\u306a\u308a\uff0c\u5927\u5e45\u306b\u5c0f\u578b\u5316\n\u304c\u5b9f\u73fe\u3067\u304d\u305f\u3053\u3068\u304c\u308f\u304b\u308b\uff0e"
        ]
    },
    {
        "image_filename": "designv11_64_0002824_chicc.2016.7554386-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002824_chicc.2016.7554386-Figure2-1.png",
        "caption": "Fig. 2: Rudder deformation fault (Right rear view)",
        "texts": [
            " \u03c4 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 I 0 0 0 0 1 2 \u03c1Sv\u03032C\u03b4r y 0 0 0 0 1 2 \u03c1Sv\u03032C\u03b4e z 0 0 1 2 \u03c1SLv\u03032m\u03b4r x 0 1 2 \u03c1SLv\u03032m\u03b4d x 0 0 1 2 \u03c1SLv\u03032m\u03b4e y 0 0 1 2 \u03c1SLv\u03032m\u03b4r z 0 0 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u23a1 \u23a2\u23a3 T \u03b4r \u03b4e \u03b4d \u23a4 \u23a5\u23a6 (3) where \u03c1, v\u0303, S, L, and T are seawater density, AUV speed, maximum cross-sectional area of AUV, AUV length and propeller thrust, respectively; \u03b4r, \u03b4e and \u03b4d represent vertical, horizontal and differential rudder angles, respectively; C\u03b4r y and C\u03b4e z describe the position derivatives derived from the corresponding force factors by \u03b4r and \u03b4e separately; m\u03b4r x , m\u03b4d x , m\u03b4e y , and m\u03b4r z are position derivatives derived from the corresponding moment factors by the relevant rudder angles. Since there are four rudders on the AUV rear, we designate them to be LR, UR, RR, and DR as shown in Fig. 2. Rudder faults could be many types, such as deformation, fracture and falling off etc., where the deformation could be much easier to occur than the other two since the latter two faults require quite a bit of external forces and the deformation is usually the first step of them. Thus we merely consider about FL and FTC against the rudder deformation faults in this paper, e.g. the deformation of ABC on RR in Fig. 2. In this section, FL and FTC against rudder deformation faults are studied through qualitative force analysis and input compensation, respectively. Two lemmas are quoted firstly to give the preconditions for the main work. Lemma 1 ([1]). If the faults contained in \u03c4 from the formula (1) are expressed as \u03c4 = \u03c4 + f (4) where \u03c4 is the desired model control input, then the faults f can be estimated by f\u0302 = x\u2212 r (5) where the vectors x and r satisfy x\u0307 = QM\u22121x\u2212QM\u22121r +QfN(\u03bd,\u03b7) +QM\u22121\u03c4 (6) r = \u2212PM\u03bd (7) Q = \u2212PM (8) Lemma 1 is used to generate the necessary fault informa- tion for FL based on the known states",
            " The accurate force analysis is able to give the detailed situations of deformation fault. But, it\u2019s difficult to analyze the deformation accurately since the deformation is undetermined. Nevertheless, the qualitative force analysis will give the localization of rudder deformation which might be able to provide enough information for the FTC. Thus, qualitative force analysis is adopted in this subsection for the FL. Consider that the deformation fault occurred on the rear corner of RR as shown in Fig. 2, where the corner has been bent downwards. We suppose that the direction of the coming flow hasn\u2019t been changed much until it reaches the bent corner of RR. Then, the qualitative force analysis could be carried out by following the hydrodynamics, which is shown in Fig. 2 where 3 forces act on the corner and the force directions are parallel to the axes of the body-fixed frame. Suppose the additive force/torque (i.e. the fault factor) vector brought in by the deformation to be f+ = [fX , fY , fZ , fK , fM , fN ]T. By analyzing the force directions and acting positions in Fig. 2, it\u2019s easy to find that the downward deformation fault of RR gives rise to fY > 0, fZ < 0, fK < 0, fM < 0, fN < 0 (10) where fX might be greater or smaller than 0 and thus isn\u2019t analyzed. Through this means, the positive/negative natures of the fault factors on the rear corners of the four rudders can be found as shown in table 1. The additive fault factors can be estimated as f\u0302 = [f\u0302X , f\u0302Y , f\u0302Z , f\u0302K , f\u0302M , f\u0302N ]T based on Lemma 1. The deformation fault can thus be localized by comparing the positive/negative natures of the estimated fault factors with the ones in table 1, where the completely matched one in the table indicates the corresponding deformation fault"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure2.29-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure2.29-1.png",
        "caption": "Fig. 2.29 Truss structure in the form of an equilateral triangle: a force boundary condition; b displacement boundary condition",
        "texts": [
            "200) The evaluation of this triple product results finally in the stiffness matrix in the global X\u2013Z coordinate system as: AE L \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 cos2 \u03b1 cos\u03b1 sin\u03b1 \u2212 cos2 \u03b1 \u2212 cos\u03b1 sin\u03b1 cos\u03b1 sin\u03b1 sin2 \u03b1 \u2212 cos\u03b1 sin\u03b1 \u2212 sin2 \u03b1 \u2212 cos2 \u03b1 \u2212 cos\u03b1 sin\u03b1 cos2 \u03b1 cos\u03b1 sin\u03b1 \u2212 cos\u03b1 sin\u03b1 \u2212 sin2 \u03b1 cos\u03b1 sin\u03b1 sin2 \u03b1 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 u1X u1Y u2X u2Y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 F1X F1Y F2X F2Y \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 . (2.201) 2.4.2 Solved Truss Problems 2.5 Example: Truss structure arranged as an equilateral triangle Given is the two-dimensional truss structure as shown in Fig. 2.29 where the trusses are arranged in the form of an equilateral triangle (all internal angles \u03b2 = 60\u25e6). The three trusses have the same length L , the same Young\u2019s modulus E , and the same cross-sectional area A. The structure is loaded by (a) a horizontal force F at node 2, (b) a prescribed displacement u at node 2. Determine for both cases \u2022 the global system of equations, \u2022 the reduced system of equations, \u2022 all nodal displacements, \u2022 all reaction forces, \u2022 the force in each rod. 2.4 Assembly of Elements to Plane Truss Structures 65 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.42-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.42-1.png",
        "caption": "Fig. 10.42 Filter inductor minor B\u2013H loop",
        "texts": [
            "41, an air gap is employed that is sufficiently large to prevent saturation of the core by the peak current I + \u0394i. The core magnetic field strength Hc(t) is related to the winding current i(t) according to Hc(t) = ni(t) c Rc Rc +Rg (10.102) where c is the magnetic path length of the core. Since Hc(t) is proportional to i(t), Hc(t) can be expressed as a large dc component Hc0 and a small superimposed ac ripple \u0394Hc, where Hc0 = nI c Rc Rc +Rg (10.103) \u0394Hc = n\u0394i c Rc Rc +Rg A sketch of B(t) vs. Hc(t) for this application is given in Fig. 10.42. This device operates with the minor B\u2013H loop illustrated. The size of the minor loop, and hence the core loss, depends on the magnitude of the inductor current ripple \u0394i. The copper loss depends on the rms inductor current ripple, essentially equal to the dc component I. Typically, the core loss can be ignored, and the design is driven by the copper loss. The maximum flux density is limited by saturation of the core. Proximity losses are negligible. Although a high-frequency ferrite material can be employed in this application, other materials having higher core losses and greater saturation flux density lead to a physically smaller device"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001407_pedes.2014.7042036-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001407_pedes.2014.7042036-Figure3-1.png",
        "caption": "Fig. 3. Space vector relationships.",
        "texts": [
            " Further, (2) can be expressed interms of ~\u03a8s & ~\u03a8r components of stator flux as: Te = 2 3 P 2 L0 \u03c3LsLr (~\u03a8s ~\u03a8r sin\u03b4) (3) From (3), it is clear that the torque developed is propotional to the cross product of ~\u03a8s & ~\u03a8r. Generally, the magnitude of the stator flux | ~\u03a8s | is kept constant, and hence the rotor flux magnitude will almost be constant and will not change instantaneously due of the large rotor time constant. Thus, in DTC, the torque developed by the machine is varied by controlling the angle \u2018\u03b4\u2019 between vectors ~\u03a8s & ~\u03a8r as shown in Fig. 3. When there is an increased torque demand, the stator flux ~\u03a8s is made to move in the anticlockwise direction by the application of voltage vectors from the look up table. This movement of ~\u03a8s will increase the angle \u03b4 and hence the torque developed. When the torque developed has to be reduced, null vectors are applied so as to stop the stator flux vector. Since rotor flux vector cannot respond as fast as stator flux vector, it continues to rotate in counter clockwise direction and thus reduces the angle \u03b4 and hence, the torque developed",
            " Every single element in Table-II denotes the space vector location used for minimizing errors in torque and flux in that sector (Fig. 4). From the stator voltage expression, d ~\u03a8s dt = ~vs\u2212~isRs (5) Generally, the voltage drop across the stator resistance is very small and can be neglected and thus it reduces to: d ~\u03a8s dt = ~vs (6) Hence, for a very small duration Ts (which is the sampling period), the stator flux follows the relation, ~\u03a8s(k+1) = ~\u03a8s(k)+~vs(k)Ts (7) where ~\u03a8s(k) & ~vs(k) denotes the value of ~\u03a8s & ~vs in the kth sampling period and, ~\u03a8s(k+1) in the (k+1)th sampling period. Fig. 3 illustrates the position of ~\u03a8s, ~\u03a8r, ~is at a given time instant. The voltage vectors that are used in look up table for error minimization are also explained here. For example, if ~\u03a8s is in sector-1 (as shown in Fig. 4), application of voltage vectors V1, V2, V3, V4, V11 and V12 will increase the flux magnitude whereas, V5, V6, V7, V8, V9 and V10 will reduce the stator flux magnitude. Likewise, voltage vectors V3, V4, V5 and V6 will increase torque in counter clockwise direction while, V9, V10, V11, and V12 will increase the developed torque in the clockwise direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure4-1.png",
        "caption": "Fig. 4. (a) Magnetic field lines (Wb/m), and (b) magnetic flux density (T) norm of the squirrel-cage induction motor. The range of contour lines is in brackects.",
        "texts": [],
        "surrounding_texts": [
            "an air gap. The stator and rotor are decomposed into disjoint163 subdomains such that \u03a9s = \u222ak\u03a9k and \u03a9r = \u222al\u03a9l hold. With164 (10) and (11), the force interactions between any of the parts \u03a9k165 and \u03a9l can be isolated and analyzed. The total force and torque166 exerted on the rotor is produced by summing up the forces and167 torques exerted by all stator parts on all rotor parts.168 In the aforementioned example, the forces between the dif-169 ferent rotor parts were not included, because they do not con-170 tribute to the total torque of the motor. Therefore, only fields171 of the sources external to the rotor must be computed. In the172 literature, the magnetic field generated by all sources external173 to any part \u03a9l is called the external magnetic field, that is,174 \u2211 k = l (Hk ,Bk ) [3].175\nIII. EDDY CURRENT PROBLEM176\nThis section presents the conventional eddy current problem,177 from which M and J necessary for decomposing the magnetic178 field are obtained. The eddy current problem concerns the field179 pairs (H,B) and (E,J) in \u03a9, where E is the electric field inten-180 sity. These satisfy Ampere\u2019s law, Gauss\u2019s law, and Faraday\u2019s law181\n\u2207\u00d7 E = \u2212\u2202tB (12)\nwhere \u2202t denotes time differentiation. The elements of the field182 pairs are coupled by the constitutive equations183\nB = \u03bcH + Br (13)\nJ = \u03c3E + Js (14)\nwhere \u03bc is the permeability, \u03c3 is the electric conductivity, Br184 is the remanence flux density, which is needed only in case185 of permanent magnets, and Js is the source current density. In186 the examples given in Section V, permeability depends on H187 making (13) a nonlinear equation.188 The Lagrangian description is used to model rotating ma-189 chines: The rotor and stator have their own frames of reference,190 fixed to the material points. The fields (H,B) and (E,J) in191 stator and rotor are to be understood in their own frames of192 reference, although no distinction is made in the notation [13].193 Again, the magnetic flux density B is expressed in terms194 of a magnetic vector potential A. Then, in order to satisfy195 (12), express E in terms of A, and in this case, it is enough196 to write197\nE = \u2212\u2202tA. (15)\nThe gradient of the electric scalar potential \u03d5 is excluded from198 (15) in the used 2-D model, because the eddy current regions199 of \u03a9 are not connected to external circuits.200 Next, (13), (6), (15), and (14) are substituted for H in Am-201 pere\u2019s law (2), which results in a parabolic PDE, whose only202 nontrivial component is203\n\u2207 \u00b7 \u03bc\u22121\u2207A \u2212 \u03c3\u2202tA + Js = \u2212(\u2207\u00d7 \u03bc\u22121Br )z . (16)\nTogether with the homogeneous Dirichlet boundary condition204 A = 0 on \u2202\u03a9 PDE (16) constitutes a BVP.205 This BVP is coupled with external circuits: The source current206 density Js depends on the voltage feed externally, and is there-207\nfore, solved within the system of equations. For more details, 208 see Appendix A. 209\nIV. IMPLEMENTATION 210\nIn this section, the BVPs presented in Sections II and III 211 are solved with the FEM. The eddy current problem is consid- 212 ered first, because its solution is necessary for computing field 213 components. 214\nThe tools used in the computations were the Gmsh finite- 215 element mesh generator [14] and the application programming 216 interface presented in [15]. 217\nA. Eddy Current Problem 218\nThe magnetic vector potential axial component A is approx- 219 imated by 220\nA \u2248 \u2211\ni\u2208I ai\u03bbi (17)\nwhere I is the index set of all mesh nodes in \u03a9\\\u2202\u03a9, and \u03bbi\u2019s and 221 ai\u2019s are the barycentric functions and the degrees of freedom, 222 respectively [16]. 223\nAccordingly, the weak Galerkin formulation of the voltage- 224 driven eddy current problem is: Find ai , i \u2208 I and Iq , q \u2208 Q 225 such that 226\n\u2211 ai\n\u222b\n\u03a9 \u03bc\u22121\u2207\u03bbi \u00b7 \u2207\u03bbj da \u2212\n\u2211 Iq\n\u222b\n\u03a9q\n\u03b2q\u03bbj da\n+ \u2211\n\u2202tai\n\u222b\n\u03a9 c\n\u03c3\u03bbi\u03bbj da = \u222b\n\u03a9m\n(\u03bc\u22121Br \u00d7\u2207\u03bbj )z da (18)\nand 227\n\u2211 \u2202tai\n\u222b\n\u03a9q\nle\u03bbi\u03b2q da + RqIq = Vq (19)\nhold for all test functions \u03bbj [17]. Here, integration by parts1 228 and the fact that A vanishes on \u2202\u03a9 have been applied. 229\nThese give a system 230\n[T ]x\u0307 + [S(x)]x = b (20)\nof nonlinear differential algebraic equations (DAE) of index 231 one [18], where the unknown coefficients are organized into a 232 vector 233\nx =\n[ a\nI\n]\n(21)\n[T ] is a singular constant-coefficient matrix, [S] is a nonsingular 234 matrix that depends on the state vector x, and b is the source 235 vector. 236\nThe backward Euler method is employed to the time 237 discretization of (20), and the resulting nonlinear algebraic sys- 238 tem of equations is solved with SUNDIALS Kinsol [19] at each 239 time step. 240\n1Leibniz rule (\u2207 \u00b7 X)Y = \u2207 \u00b7 (XY ) \u2212 X \u00b7 \u2207Y with Gauss\u2019s divergence theorem, and Leibniz rule (\u2207\u00d7 X)Y = \u2207\u00d7 (XY ) \u2212 X \u00d7\u2207Y with Stokes\u2019 theorem, respectively.",
            "B. Magnetic Field Component241\nThe field Ak is approximated with \u2211 i\u2208I ak i \u03bbi . Then, the weak242 Galerkin formulation of the field component problem is: Find243\nak i , i \u2208 I such that244\n\u2211 ak\ni\n\u222b\n\u03a9 (\u2207 \u00b7 \u03bc\u22121\n0 \u2207\u03bbi)\u03bbj da = \u2212 \u222b\n\u03a9k\n(\u2207\u00d7 Mk )z\u03bbj da\n\u2212 \u222b\n\u03a9k\nJk\u03bbj da (22)\nholds for all test functions \u03bbj . Integrate by parts and use the fact245 that Ak vanishes on \u2202\u03a9 to obtain246\n\u2211 ak\ni\n\u222b\n\u03a9 \u03bc\u22121\n0 \u2207\u03bbi \u00b7 \u2207\u03bbj da = \u2212 \u222b\n\u03a9k\n(Mk \u00d7\u2207\u03bbj )z da\n+ \u222b\n\u03a9k\nJk\u03bbj da \u2200j. (23)\nThis corresponds to a system247\n[Sk ]ak = bk (24)\nof linear algebraic equations, where the degrees of freedom are248 organized into a vector ak , and the elements of the system matrix249 [Sk ] and the input bk vector are250\nSk ji =\n\u222b\n\u03a9 \u03bc\u22121 0 \u2207\u03bbi \u00b7 \u2207\u03bbj da (25)\nbk j = \u2212\n\u222b\n\u03a9k\n(Mk \u00d7\u2207\u03bbj )z da + \u222b\n\u03a9k\nJk\u03bbj da. (26)\nThe linear system matrix [Sk ] is the same for all k; hence,251 a single matrix decomposition enables the computation of all252 pairs (Hk ,Bk ).253\nV. EXAMPLES254\nIn this section, decomposition of the magnetic field and torque255 is demonstrated with a squirrel-cage induction motor and a per-256 manent magnet synchronous motor.257\nA. Squirrel-Cage Induction Motor258\nFig. 2 shows a quarter of the model geometry of the squirrel-259 cage induction motor. The essential motor parameters are in260 Table I, and the BH-curve of the iron cores and shaft are in261 Fig. 3. The iron cores are assumed nonconducting, whereas262 the shaft and squirrel-cage have otherwise cite it at appropriate263 place. a conductivity of 4.3e6 S/m and 3.2e7 S/m, respectively.264 Each stator coil is connected to a balanced three-phase voltage265 Vq = V0 cos(2\u03c0ft + q2\u03c0/3), where f = 50 Hz, q \u2208 {0, 1, 2}266 and V0 = 400 V. Rotation frequency is fixed to 24.5 Hz, which267 results in a 2% slip in a four-pole motor.268 The motor fields are first solved from the eddy current prob-269 lem, and then, the magnetic field is decomposed into four com-270 ponents: the stator iron field, the stator coil current field, the271 rotor iron field, and the rotor induction current field. The field272 lines of the magnetic field and its components are presented in273 Figs. 4 and 5, respectively. In this example, the computation of274 each magnetic field component [assembly and solution of the275",
            "linear system (24)] took less than 4% of the solution time of the276 motor magnetic field [assembly and solution of the nonlinear277 DAE (20)].278\nThe difference between B and \u2211\nk Bk should theoretically be279 zero. It was computed for FEM approximations, and the relative280 two-norm281 \u221a \u221a \u221a \u221a \u222b \u03a9 \u2016B \u2212 \u2211 k Bk\u20162 da\n\u222b \u03a9 \u2016B\u20162 da\n(27)\nremained below 5.3e-10, and the difference between the com-282\nputed A and \u2211\nk Ak remained below 7.5e-12 Wb/m for the283 studied time span.284 The stator field components are necessary for calculating the285 torque exerted on the rotor. In that case, applying (11) to the286 rotor iron and bars results in a total of four torque components:287 stator iron to rotor iron, stator iron to rotor current, stator current288 to rotor iron, and stator current to rotor current.289 Fig. 6 presents the total torque on the rotor computed with290 the MST and the sum of all torque components on the rotor.291 In this example, there is approximately a 1.5% difference be-292 tween the computed total rms torques. We have observed that293 the difference decreases with a refined mesh. This is consistent294 with the results presented in [20], where forces obtained with295 the MST and ESM (using external field) are compared with dif-296 ferent mesh densities (the sum of stator field components is the297 external field for the rotor).298 It is evident from Fig. 7 that a major part of the torque ripple299 produced by the squirrel-cage induction motor is due to stator300 iron to rotor iron interaction. In this example, there is also a301\nnotable phase shift between the stator iron to rotor iron torque 302 component and the other torque components. 303\nB. Permanent Magnet Motor 304\nFig. 8 presents the model geometry of the permanent magnet 305 motor and Table II the essential motor parameters. The rotor has 306 four surface-mounted magnets with 1-T remanence flux density 307 and 1.05 \u03bc0 permeability. The BH-curve of the iron cores and 308 shaft are presented in Fig. 3. The iron cores are assumed non- 309 conducting, whereas the shaft and magnets have a conductivity 310 of 4.3e6 S/m and 6.7e5 S/m, respectively. 311\nThe permanent magnet motor fields are decomposed into four 312 components: the stator iron field, the stator coil current field, the 313 rotor iron field, and the rotor magnets field. The magnetic field 314 and its components are presented in Figs. 9 and 10, respectively. 315\nThe relative two-norm (27) was used to measure the differ- 316\nence between B and \u2211\nk Bk . This remained below 4.8e-10, and 317\nthe absolute difference between A and \u2211\nk Ak remained below 318\n7.9e-12 Wb/m for the studied time span. For comparison, in [8], 319 the FR method was applied to a permanent magnet motor, and 320 there was approximately a 5% error in the magnetic flux density 321 in the motor air gap. 322"
        ]
    },
    {
        "image_filename": "designv11_64_0000973_holm.2014.7031019-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000973_holm.2014.7031019-Figure7-1.png",
        "caption": "Fig. 7: Macroscopic and microscopic illustration: a) of the electrical contact; b) of the Lorenz force developing in two opposing conductors and their magnetic fields",
        "texts": [
            " R R = 2\u00b7 k \u00b7F + F \u00b7H (2) with k = 3 4 1- E + 1- E \u00b7 1 r - 1 r (3) In the stationary electrical contacts, the contact force FC is composed of the mechanical applied load (pressure finger force F ), electrodynamics forces (Lorenz force F ) that depends on the contact area A and the amplitude of the brush current [8, 2]. In our case we have also the acceleration force F due to the circumferential speed v and the eccentricity of the slip ring surface (7). All the forces that appear in our systems are shown in Fig. 6. By narrowing of contact region, the current flow paths along the boundary of the contact area are parallel to each other (Fig. 7a). Their own magnetic fields appear opposite to each other. The reason is the small contact surfaces in the contact region. The occurred electromagnetic forces FA are opposite so that the two bodies push each other (Fig. 7b). The contact force F is showed in (5). The used variables in (5), (6) and (7) are defined in the nomenclature. F =F - F +F +F (5) F = 8 \u00b7i ln 8 \u00b7H\u00b7A \u00b7i (6) F =m \u00b7 4 dh v 60 D (7) III. DESCRIPTION OF INFLUENCE PARAMETER ON SPARKING APPEARANCE With the developed electrical system in [9], the influences of the contact force F , the circumferential speed and the dependence on the environmental influences on the current flow and the contact voltage can be measured. These measurements help us to have a global view on the appearance of mirco- or macro brush fires, but it is still difficult to characterize this phenomenon"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003707_978-3-540-36045-2_3-Figure3.13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003707_978-3-540-36045-2_3-Figure3.13-1.png",
        "caption": "Fig. 3.13 Double wishbone wheel suspension",
        "texts": [
            " The five-link suspension The five-link suspension (in German \u2018\u2018Raumlenkerachse\u2019\u2019) embodies a closed kinematic chain with a very high degree of inner coupling of the inherent kinematic loops (Fig. 3.12). Because in the assignment of degrees of freedom of the system, the isolated rotation of the suspension arms around their longitudinal axes are neglected, the spherical joints at the chassis end are modeled as CARDAN joints, without loss of generality. Double wishbone wheel suspension In contrary to the five-link wheel suspension, the double wishbone suspension (Fig. 3.13) possesses only a weak kinematic coupling between the inherent two kinematic loops. Thus the associated kinematic analysis can be stated and solved explicitly, as shown in Sect. 3.5.5 and Chap. 6. The position of all the bodies in a kinematic chain with f total degrees of freedom can uniquely be defined through f independent coordinates, the generalized coordinates. Examples Kinematic chains with tree structure (planar) In Fig. 3.14 the natural joint coordinates bi (revolute joints) can directly be selected as the generalized coordinates"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000876_ipec.2014.6869589-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000876_ipec.2014.6869589-Figure1-1.png",
        "caption": "Fig. 1. Rotor configuration",
        "texts": [
            " This paper deals with two types of PMs to imitate the demagnetization. One is a situation that the PM of one pole is reduced in the radial direction. The other is a situation that the PM of one pole is reduced in the axial direction. The volume of one of four PMs is reduced by 10%, 20 % and 30 %, namely, the amount of 2.S %, S % and 7.S % of the PM is decreased. This paper clarifies the effect of the demagnetization of PMs on the performance of PMSMs under controlled by both a V/f constant and vector strategy. II. PMSM WITH PERMANENT MAGNET DEFECT Fig. 1 shows the rotor configuration of the experimental PMSM. This motor is a l.S-kW, 3000-min-\\ 4.8-N\u00b7m, S.6-A, four-pole machine. One of four poles is demagnetized. Fig. 2 shows the demagnetization in radial direction, we call this situation radial demagnetization hereafter. One pole is composed of four PMs, and the thickness of four PMs which compose one pole is reduced by 10 %, 20 % and 30 % in order to imitate the demagnetization or imperfect magnetization. Fig. 3 shows the demagnetization in axial direction, we call this situation axial demagnetization hereafter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.25-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.25-1.png",
        "caption": "FIGURE 8.25",
        "texts": [
            " As shown in Figure 8.24, a 150 lb external force pointing upward was applied on the road profile cam to mimic the wheel load due to racecar weight (445 lb) and driver weight (155 lb). An equilibrium analysis was first carried out. The equilibrium state of the racecar was assumed as the initial condition for the dynamic simulation, in which the racecar started in equilibrium on the flat road and then reached the first hump. A spring and a damper were also defined in the dynamic analysis, as shown in Figure 8.25. The physical position of the spring is shown in Figure 8.26. The spring rate and the damping coefficient were 100 lb/in. and 10 lb/(in./sec), respectively. The free length of the spring was 5.5 in. Note that when the shock was fully extended to its maximum length, the spring length was 4 in., which implies a 150 lb preload. The dynamic simulation (Case A) was carried out assuming a racecar speed of 4.74 mph. The shock travel is shown in Figure 8.27. Note that the shock length was allowed to vary between 7"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000099_978-94-007-7194-9_14-1-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000099_978-94-007-7194-9_14-1-Figure4-1.png",
        "caption": "Fig. 4 Two DOF cylindrospheric module used in the neck, shoulder, and elbow of NAO",
        "texts": [
            " In addition, this structure helps better distribute the power between the hip roll joint and the pelvis joint and creates a specific motion style to the NAO humanoid. If the very first prototypes of NAO were based on servo motors; the integration of these actuators would not be compatible with the very strong constraints that Aldebaran aimed for on the physical appearance of the robot. For this reason, specific actuators, based on DC motors, gears, and position sensors were created. 10 degrees of freedom in NAO rely on the same type of actuation module called the cylindrospheric module (Fig. 4). This module includes two motors that provide two perpendicular motions. This module is very compact and makes the integration of motorization very easy. This module is used in the shoulder, at the elbow, and in the neck of NAO. This modularity is an important aspect to consider when regarding the goal of mass production for NAO. Another constraint that led to the design of proprietary actuators was research in backdrivability. Classically, humanoid robots use harmonic drives which provide high reduction ratio in a small volume, but are not backdrivable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003415_ecce.2016.7854897-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003415_ecce.2016.7854897-Figure8-1.png",
        "caption": "Fig. 8. Schematic diagram of current signal and vibration signal acquisitions from an Air Breeze wind turbine located in a wind tunnel.",
        "texts": [
            " 7(c) shows the order-domain spectrum of the resampled current envelope signal obtained by the angular resampling method. This spectrum still contains some interfering peaks around order 15 (marked by an ellipse) and, thus, is inferior to the MFS in Fig. 7(b) for the bearing fault diagnosis. B. Vibration Signal Analysis In this subsection, the vibration signal of a 160 W Air Breeze wind turbine is used for bearing fault diagnosis by the proposed method. The schematic diagram of the current and vibration signal acquisition is illustrated in Fig. 8. The wind turbine is located in a wind tunnel with controllable wind flows to simulate the real-world working conditions. The range of the wind speed varies from 0 to 10 m/s. The number of pole pairs of the PMSG is 6. One of the bearings supporting the main shaft is taken for testing. An accelerometer is mounted on the casing of the wind turbine to measure the vibration. One phase stator current signal of the PMSG is measured by a Fluke 80i110s AC/DC current clamp. The sampling frequency of these two signals is 10 kHz"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002184_ecce.2014.6954112-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002184_ecce.2014.6954112-Figure1-1.png",
        "caption": "Fig. 1. Block diagram of existing high switching freq. stator current observer",
        "texts": [
            " Therefore, observers are essential to implement DB-DTFC. In this section, the discrete time model of the existing stator current observer is first introduced. In order to form a simple transfer function for the stator current observer model, the cross-coupling terms are assumed to be perfectly decoupled. The cross-coupling terms are added to the system as feedforward signals to the system, as depicted in the block diagram of the existing high switching frequency stator current observer shown in Fig. 1. The discrete time voltage model of the existing stator flux linkage observer is formed with the stator voltage as the latched input, the stator current as a ramp signal and with the cross-coupled terms decoupled. The discrete time model of the Gopinath-style stator flux linkage observer composed of discrete time current and voltage models is developed and shown in Fig. 2. pm\u03bb + + + \u2212 + \u2212++ + sr \u2227 11 \u2212\u2212 z TK fio 1\u2212z Ts foK + )1(2 )1( \u2212 + z zTs \u2212 ( ( ) ( ))r r r qs dsz j z\u03c9 \u03bb \u03bb\u2212 sL \u2227 ( )r dqsi z * ( )r dqs z\u03bb \u2227 ( )r dqsv z ( )r zdqs\u03bb \u2227 Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure3-1.png",
        "caption": "Figure 3 Control of direction, (a) straight position (b) turn (see online version for colours)",
        "texts": [
            " Figure 2(a) shows a passive manipulator, and Figure 2(b) shows the proposed duplex mechanism. The duplex mechanism consists of two passive manipulators. Two wires are installed along either side of each manipulator. Every passive joint of the manipulator has a locking mechanism, and a hose for the locking mechanisms is installed along the centre of the manipulator. A rail is installed on each manipulator. The two manipulators are connected through the rails and can move along each other via the rails [Figure 2(b)]. As shown in Figure 3, the head link of each manipulator can be turned by pulling either wire. As shown in Figure 4, a piston is attached to the central hose. Each manipulator has its own piston. By pushing on the piston, the hose expands until the two friction materials are engaged. The joint is thus locked (Figure 5). The proposed duplex mechanism enables movement as follows (Figure 6). First, at the start of a curve, one manipulator is moved forward. Next, the direction of the head of the manipulator is changed by pulling a wire"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002342_indicon.2015.7443272-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002342_indicon.2015.7443272-Figure2-1.png",
        "caption": "Figure 2: Yaw Plane Dynamics",
        "texts": [
            " Based on the Newton's and Euler's equations the six degrees of freedom motion equation for an AUV can be in written in terms of body fixed coordinates: m[u - vr + wq - Xg(qZ + rZ) + Yg(pq - i\") + zg(pr + q) = I Xext m[v - wp + ur - yg(rZ + pZ) + Zg(qr - v) + Xg(qp + i\") = I Yext m[w - uq + vp - Zg(pZ + qZ) + Xg(rp - q) + yg(rq + v) = I Zext IxxV + (Izz - Iyy)qr + m[Yg(w - uq + vp) - Zg(V - wp + ur)] = I Kext IyyiJ + (Ixx - Izz)rp + m[Zg(u - vr + wq) - Xg(W - uq + vp)] = I Mext Izzi\" + (Iyy - Ixx)pq + m[Xg(v - wp + ur) - Yg(U - vr + wq)] = I Next (3) Reduced order modelfor yaw dynamics In this paper we proposed the control design for yaw plane dynamics. For yaw plane maneuvering yaw angle has to be controlled. As shown in Fig. 2, when a rotation of angle tp is given along the Z axis, the X and Y axes are shifted to the new position Xl and Ylrespectively. Therefore we can neglect the out of plane terms to derive the following three term state vector as given in [2] which describe only the yaw plane dynamics. The linearized equations of motion for yaw plane are represented in matrix form where the vehicle has steady-state speed [1]. [ (m - Yv) (mxG - Vi) \ufffdml+ (mxG 0 - Nv) (Izz - Ni) 0 [-Yv (mU - Yr) \ufffdl [\ufffdH\ufffdl 0, -Nv (mxGU - Nr) 0 -1 (4) The transfer function for yaw rate 'r' in terms of rudder angle orcan be found by substituting the value of system parameters given in TABLE 1 and by neglecting the out of plane terms from the above three term state vector"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000497_9781119011804.ch6-Figure6.7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000497_9781119011804.ch6-Figure6.7-1.png",
        "caption": "FIGURE 6.7 BioPhotonics workstation equipped with fluorescence system and temperature control [35]. RSI. The heating stage lies between the opposing objectives that deliver counterpropagating beam traps. Fluorescence is imaged by the top objective through the dichroic mirror and filters to the CCD camera. Brightfield imaging uses white light illumination through the bottom objective.",
        "texts": [
            " Optical trapping systems have been combined with different imaging modalities such as brightfield imaging, phase contrast, and digital holographic microscopy [32], which can also retrieve quantitative phase information [33, 34]. Optical trapping systems have also been combined with spectroscopic imaging, some of which will be discussed below. The need for sensitive monitoring of trapped cells may be addressed by integrating various instrumentation for realizing environmental controls and sensors, where the specific instrumentation depends on the cells under study. One such Biophotonics workstation is schematically depicted in Figure 6.7. This system integrates the capacity for simultaneous optical manipulation, heat stress, fluorescence, and intracellular pH measurements [35]. Its multifunctionality can be utilized, for example, in yeast cell studies where the measured fluorescence enables the calculation of intracellular pH distribution, which turns out to be a good indicator of yeast cell health [36]. As mentioned previously, subcellular structures influence the scattering properties of the cell. Hence, analysis of the scattered trapping beam itself can yield information on the internal dynamics of trapped cell, for example, to monitor cellular processes [37]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000849_j.mechmachtheory.2014.05.001-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000849_j.mechmachtheory.2014.05.001-Figure10-1.png",
        "caption": "Fig. 10. New second-type of Twisting Tower derived from the first basic Twisting Tower.",
        "texts": [
            " 9 shows two new types of Twisting Tower derived from the third basic Twisting Tower in three different positions. In both cases (n = 5 and n = 12) for p/q = 2/1 have been chosen. According to the results presented in formula (10), in the kinematic chains connecting the top body with the body at the bottom, the values of the parameters pi and qi are to be chosen arbitrarily but equally and the angles \u03b1i are arbitrary in the successive fields shown in Fig. 5. The number n of kinematic chains surrounding the vertical axes is arbitrary as well. Dates of the Twisting Towers in Fig. 10. n = 3:r = 100 mm, \u03b11 = 70\u00b0,\u03b13 = 100\u00b0,\u03b12 = 58.52\u00b0:n = 6: r = 100 mm,\u03b11 = 70\u00b0, \u03b13 = 100\u00b0, \u03b12 65.05\u00b0. Dates of the Twisting Towers in Fig. 11. n = 4:r = 100 mm, \u03b11 = 25\u00b0, \u03b12 = 60\u00b0, \u03b13 = 46\u00b0,\u03b14 = 91.73\u00b0; n = 7:r = 100 mm, \u03b11 = 17\u00b0, \u03b12 = 60\u00b0, \u03b13 = 30\u00b0, \u03b14 = 88.97\u00b0. Dates of the Twisting Towers in Fig. 12. n = 5:r = 100 mm, \u03b11 = 28.65\u00b0, \u03b12 = 22.92\u00b0, \u03b13 = \u03b11, \u03b14 = 57.67\u00b0, \u03b15 = 40.25\u00b0, \u03b16 = 85.94\u00b0; n = 12: r = 100 mm,\u03b11 = 28.65\u00b0, \u03b12 = 14.18\u00b0, \u03b13 = \u03b11, \u03b14 = 48.10\u00b0, \u03b15 = 25.10\u00b0, \u03b16 = 85.94\u00b0. This paper first presents three basic Twisting Towers which are derived from three special Archimedean polyhedrons by omitting square faces and connecting the remaining faces via double rotor joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000978_pic.2014.6972391-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000978_pic.2014.6972391-Figure2-1.png",
        "caption": "Fig. 2. Situation space diagram",
        "texts": [
            " An intrinsic attribute is an attribute which negatively affects the inference of a situation if its value is not within the accepted region of values. An optional attribute is an attribute that assists in inferring a situation, but sensing values outside the accepted region would not weaken the support for getting the situation. A context state describes the current situation of concerned context-aware application, denoted by vector iS . iS is a collection of N context attribute values that are used to represent a specific situation at a time t . Therefore, a context state can be denoted by  t N ttt i aaaS ,,, 21  . As shown in Fig. 2, situation space iR is a collection of ranges corresponding to some predefined situation (e.g., Sickness or Normal behavior), and it can be denoted by  ** 2 * 1 ,,, R N RR i aaaR  . An acceptable range *R ia is defined as a set of elements that satisfies a predicate *R . S and S describe two different context states respectively. S is an element inside the region iR and S is an element outside the region iR . xa , ya , za describe three different context attributes respectively, so the acceptable ranges of these three attributes make up the situation space iR "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001901_s1068798x15080055-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001901_s1068798x15080055-Figure2-1.png",
        "caption": "Fig. 2. Cross sections of the RZA S 2000 (a) and RZAM S 1000 (b) gear systems.",
        "texts": [
            "774 External diameter of spiroid screw da1, mm 42 48 Internal/external diameter of spiroid gear di2/de2, mm 138/175 120/155 Maximum torque on gear\u2019s output shaft T2max, N m 2000 1000 Limiting static torque on output shaft T2li, N m 4000 2000 Mass m, kg 17 11 The gears (Fig. 1) include a housing 1 that contains the spiroid screw 6; input flange (lid) 3; adapter 4 (the output shaft for the quarter turn gear) or a protective dome (for the multiturn gear); and a transmission consisting of the spiroid screw 6 rotating in roller bear ings, and the gear 7, rotating in slip bands. Auxiliary com ponents include indicators, speed limiters, and structural components. The RZA S 2000 system (Fig. 2a) is based on spiroid gear 1, radial bearings 2, and an end bearing 3, while the RZAM S 1000 system (Fig. 2b) is based on gear 1, radial bearings 2, and two end bearings 3. The following slip bearings are chosen: 1. Metal\u2013fluoroplastic bearings manufactured from a strip consisting of a steel base (low carbon steel) and a layer (0.20\u20130.35 mm) of sintered bronze powder that has been infiltrated and coated with a layer (0.01\u20130.04 mm) of polytetrafluoroethylene and molybdenum disulfide. The main downside of this type is high cost [4]. These bearings withstand heavy loads, operate without lubricant over a broad temper ature range, and maintain a low frictional coefficient"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure5-1.png",
        "caption": "Fig. 5. Magnetic field components of the squirrel-cage induction motor: (a) stator iron field lines, (b) stator current field lines, (c) rotor iron field lines, (d) rotor current field lines. The range of the contour lines is in brackets in Wb/m.",
        "texts": [],
        "surrounding_texts": [
            "B. Magnetic Field Component241\nThe field Ak is approximated with \u2211 i\u2208I ak i \u03bbi . Then, the weak242 Galerkin formulation of the field component problem is: Find243\nak i , i \u2208 I such that244\n\u2211 ak\ni\n\u222b\n\u03a9 (\u2207 \u00b7 \u03bc\u22121\n0 \u2207\u03bbi)\u03bbj da = \u2212 \u222b\n\u03a9k\n(\u2207\u00d7 Mk )z\u03bbj da\n\u2212 \u222b\n\u03a9k\nJk\u03bbj da (22)\nholds for all test functions \u03bbj . Integrate by parts and use the fact245 that Ak vanishes on \u2202\u03a9 to obtain246\n\u2211 ak\ni\n\u222b\n\u03a9 \u03bc\u22121\n0 \u2207\u03bbi \u00b7 \u2207\u03bbj da = \u2212 \u222b\n\u03a9k\n(Mk \u00d7\u2207\u03bbj )z da\n+ \u222b\n\u03a9k\nJk\u03bbj da \u2200j. (23)\nThis corresponds to a system247\n[Sk ]ak = bk (24)\nof linear algebraic equations, where the degrees of freedom are248 organized into a vector ak , and the elements of the system matrix249 [Sk ] and the input bk vector are250\nSk ji =\n\u222b\n\u03a9 \u03bc\u22121 0 \u2207\u03bbi \u00b7 \u2207\u03bbj da (25)\nbk j = \u2212\n\u222b\n\u03a9k\n(Mk \u00d7\u2207\u03bbj )z da + \u222b\n\u03a9k\nJk\u03bbj da. (26)\nThe linear system matrix [Sk ] is the same for all k; hence,251 a single matrix decomposition enables the computation of all252 pairs (Hk ,Bk ).253\nV. EXAMPLES254\nIn this section, decomposition of the magnetic field and torque255 is demonstrated with a squirrel-cage induction motor and a per-256 manent magnet synchronous motor.257\nA. Squirrel-Cage Induction Motor258\nFig. 2 shows a quarter of the model geometry of the squirrel-259 cage induction motor. The essential motor parameters are in260 Table I, and the BH-curve of the iron cores and shaft are in261 Fig. 3. The iron cores are assumed nonconducting, whereas262 the shaft and squirrel-cage have otherwise cite it at appropriate263 place. a conductivity of 4.3e6 S/m and 3.2e7 S/m, respectively.264 Each stator coil is connected to a balanced three-phase voltage265 Vq = V0 cos(2\u03c0ft + q2\u03c0/3), where f = 50 Hz, q \u2208 {0, 1, 2}266 and V0 = 400 V. Rotation frequency is fixed to 24.5 Hz, which267 results in a 2% slip in a four-pole motor.268 The motor fields are first solved from the eddy current prob-269 lem, and then, the magnetic field is decomposed into four com-270 ponents: the stator iron field, the stator coil current field, the271 rotor iron field, and the rotor induction current field. The field272 lines of the magnetic field and its components are presented in273 Figs. 4 and 5, respectively. In this example, the computation of274 each magnetic field component [assembly and solution of the275",
            "linear system (24)] took less than 4% of the solution time of the276 motor magnetic field [assembly and solution of the nonlinear277 DAE (20)].278\nThe difference between B and \u2211\nk Bk should theoretically be279 zero. It was computed for FEM approximations, and the relative280 two-norm281 \u221a \u221a \u221a \u221a \u222b \u03a9 \u2016B \u2212 \u2211 k Bk\u20162 da\n\u222b \u03a9 \u2016B\u20162 da\n(27)\nremained below 5.3e-10, and the difference between the com-282\nputed A and \u2211\nk Ak remained below 7.5e-12 Wb/m for the283 studied time span.284 The stator field components are necessary for calculating the285 torque exerted on the rotor. In that case, applying (11) to the286 rotor iron and bars results in a total of four torque components:287 stator iron to rotor iron, stator iron to rotor current, stator current288 to rotor iron, and stator current to rotor current.289 Fig. 6 presents the total torque on the rotor computed with290 the MST and the sum of all torque components on the rotor.291 In this example, there is approximately a 1.5% difference be-292 tween the computed total rms torques. We have observed that293 the difference decreases with a refined mesh. This is consistent294 with the results presented in [20], where forces obtained with295 the MST and ESM (using external field) are compared with dif-296 ferent mesh densities (the sum of stator field components is the297 external field for the rotor).298 It is evident from Fig. 7 that a major part of the torque ripple299 produced by the squirrel-cage induction motor is due to stator300 iron to rotor iron interaction. In this example, there is also a301\nnotable phase shift between the stator iron to rotor iron torque 302 component and the other torque components. 303\nB. Permanent Magnet Motor 304\nFig. 8 presents the model geometry of the permanent magnet 305 motor and Table II the essential motor parameters. The rotor has 306 four surface-mounted magnets with 1-T remanence flux density 307 and 1.05 \u03bc0 permeability. The BH-curve of the iron cores and 308 shaft are presented in Fig. 3. The iron cores are assumed non- 309 conducting, whereas the shaft and magnets have a conductivity 310 of 4.3e6 S/m and 6.7e5 S/m, respectively. 311\nThe permanent magnet motor fields are decomposed into four 312 components: the stator iron field, the stator coil current field, the 313 rotor iron field, and the rotor magnets field. The magnetic field 314 and its components are presented in Figs. 9 and 10, respectively. 315\nThe relative two-norm (27) was used to measure the differ- 316\nence between B and \u2211\nk Bk . This remained below 4.8e-10, and 317\nthe absolute difference between A and \u2211\nk Ak remained below 318\n7.9e-12 Wb/m for the studied time span. For comparison, in [8], 319 the FR method was applied to a permanent magnet motor, and 320 there was approximately a 5% error in the magnetic flux density 321 in the motor air gap. 322",
            "Fig. 11 compares the total torques exerted on the rotor calcu-323 lated with the MST and with field decomposition. There is ap-324 proximately a 0.2% difference in the computed total rms torques.325 Fig. 12 presents the torques exerted by the stator iron and sta-326 tor coil current on the rotor iron and rotor magnets, respectively.327 It is evident from Fig. 12 that most of the permanent magnet328 motor torque and torque ripple is produced by stator iron to rotor329 magnets interaction.330\nVI. DISCUSSION331\nIn the examples in Section V, the motor magnetic fields were332 decomposed according to iron cores, current-carrying conduc-333\ntors, and permanent magnets. It is possible to decompose the 334 magnetic field in many other ways as well. For example, one 335 could refine, if necessary, the field generated by the stator core 336 into the field components of stator teeth and the rest of the sta- 337 tor iron. Furthermore, the field caused by stator coils can be 338 divided into field components caused by separate phases. The 339 refinement of field components also results in a number of new 340 torque components. 341\nThe results give new quantitative information on the parts of 342 the motor that produce torque. Such insights are useful in, for 343 example, minimizing the torque ripple. For example, the effect 344 of changes made to the shape of the stator tooth, the perma- 345 nent magnet, the squirrel-cage bar, winding distribution, and 346 the feeding current waveform can be seen in the corresponding 347"
        ]
    },
    {
        "image_filename": "designv11_64_0001850_amm.770.402-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001850_amm.770.402-Figure2-1.png",
        "caption": "Fig. 2. Mises stress,\u041cP\u0430 (\u0430) and contact stress,\u041cP\u0430 (b) for lipGOST 14896-84 at working fluid pressureP=50 \u041cP\u0430",
        "texts": [],
        "surrounding_texts": [
            "Operating capacity of power cylinders, including hydraulic jacks and hydraulic legs of power supports, is determined according to the capacity of the leap seal in piston \u2013 working cylinder clearance, as well as to the sealed clearance size. This paper provides a comparative operation analysis of three seals, manufactured according to GOST 6678-72, GOST 14896-84 and GOST 6969-54; they differ in form and geometrical parameters."
        ]
    },
    {
        "image_filename": "designv11_64_0002827_jae-162095-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002827_jae-162095-Figure2-1.png",
        "caption": "Fig. 2. Photo of the magnetic gear (Type A and B).",
        "texts": [
            "jp. 1383-5416/16/$35.00 c\u00a9 2016 \u2013 IOS Press and the authors. All rights reserved Figure 1 shows the proposed magnetic gear with the multiple layers of the inner and the outer rotors. The ring magnets and the iron plates with gear-like shape are stacked in both rotors. These magnets are magnetized along the axial direction and the same poles face each other across the iron plates to concentrate the magnetic flux. The center steel segment is made of electromagnetic irons and nonmagnetic material. Figure 2 shows photos of parts of the proposed magnetic gear Type A. As shown in Fig. 2, the magnetic gear is structured of (a) an outer iron plate of Type A, (b) an outer ring magnet, (c) an inner iron plate of Type A, and (d) an inner ring magnet. Also an assembled inner rotor and an assembled center rotor are (e) and (f) respectively. As the Type B, another inner and outer rotors are shown in Figs 2(g) and (h). The rotors of Type B has iron plates with the neodymium magnets magnetized along radial direction between the teeth. Table 1 shows the dimensions of the prototype magnetic gear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001478_s12206-014-0314-0-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001478_s12206-014-0314-0-Figure5-1.png",
        "caption": "Fig. 5. Regions from which the experimental data were collected.",
        "texts": [
            " On the front region, xf = Y and yf = -X. On the upper side, xu = -X and yu = -Y. Based on the coordinate system, the contact length on every side was taken as the abscissa, with the center of the contact length being the origin of the local coordinate system. The X-ring has a very complicated geometry. Careful analysis of the the internal stresses using the photoelastic experimental hybrid method can only adequately be achieved by sub-dividing the specimen in to six regions for easier analysis. Fig. 5 shows the various regions from which the experimental data were collected. The regions with higher stresses provide useful information needed for design purposes. The six regions are the lower left region (region 1), lower right region (region 2), the bottom front region (region 3), the top front region (region 4), the upper right region (region 5), and the upper left region (region 6). The isochromatic fringe patterns and internal stress contours of each section were rotated or not rotated by taking the contact line to coincide with the abscissa (x-axis)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000669_sice.2015.7285529-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000669_sice.2015.7285529-Figure5-1.png",
        "caption": "Fig. 5 AR.Drone2.0. Fig. 6 Prime17W",
        "texts": [
            "030m and 0.001m. Furthermore, in Fig. 10(d),10(e), we can confirm the velocity of the return path is suppressed. To consider the disturbance database, it is possible for the proposed method to use the disturbance information in advance. We can also confirm the effectiveness in the Fig. 10(a) that the UAV takes the preliminary action before it is disturbed. We use the Parrot AR.Drone which is a commercially available quadcopter. The appearance and the definition of the attitude angle is depicted in Fig. 5. AR.Drone is equipped with various sensors. The internal sensors are 3-axis gyroscope, an 3-axis accelerometer, an 3-axis magnetic sensor and a barometer. The external sensors are altitude measurement ultrasonic sensor and two cameras. Further, we use the three-dimensional measurement cameras Prime17W depicted in Fig. 6 to estimate the selfposition. The frame rate of Prime17W is 360 fps. The position of the object can be estimated with an error of less than \u00b10.1mm per 2.8ms. Block diagram of the entire experimental system using these devices are shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002394_j.jestch.2016.05.003-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002394_j.jestch.2016.05.003-Figure3-1.png",
        "caption": "Fig. 3. Test bearings (a) 32205BJ2/Q (b) 30305C, and (c) defect creation using EDM.",
        "texts": [
            " The validation experiment is a final step in verifying the conclusions from the previous round of experimentation. In order to validate the results obtained from Eq. (14), confirmation experiments were conducted for obtaining the response characteristics in which a selected number of tests are run under specified conditions as listed in Table 7. The specifications of the test bearing for solution of the Eq. (14) are taken from Table 8. These bearings are damaged by inducing artificial spalls of different sizes to the rollers using electric discharge machining as shown in Fig. 3(a)\u2013(c). The characteristic vibration defect frequencies of the test bearings are to be calculated using the expressions as given in Table 9 [56]. The results of the validation experiments are listed in Table 10. For the validation test 1, the experimental set-up is fitted with roller damaged 30305C taper roller bearing and the rotor shaft speed was set to 900 rpm. Fig. 4 shows the experimental spectrum obtained during this test 1. The significant peak of 0.283 m/s2 is observed at the 487 Hz frequency which is nearly equal to the 6th harmonics of the theoretical roller defect frequency of 76"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000288_978-3-642-40849-6-Figure2.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000288_978-3-642-40849-6-Figure2.2-1.png",
        "caption": "Fig. 2.2. Equivalent model of thrusting system of shield machine",
        "texts": [
            " 5, the angular velocity of cutterhead can be rewritten as: (6) By differentiating p\u03c9 with respect to time, the angular acceleration is given by: ' xx T T p bp p bp bp yy bp zz I I I    = =      I A I A A A ( ) 16 1 i i p p p p p e d p i= + \u00d7 = + \u00d7I \u03c9 \u03c9 I \u03c9 M F r cos cos cos cos sin sin cos 0 sin cos sin sin cos y z y z y bp z z y z y z y \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8  \u2212  =    \u2212  A 0 T p bp y z\u03b8 \u03b8 =  \u03c9 A L   1 0 sin 0 0 1 0 0 0 cos y p y y z \u03b8 \u03b8 \u03b8 \u03b8        =           \u03c9   Dynamical Behavior of Redundant Thrusting Mechanical System in Shield Machines 693 (7) Substitute Eq.4, 6, 7 into Eq. 3, and use the basis vector 2,3 0 1 0 0 0 1   =     e , the system Euler formulation can be rewritten as: (8) Combine Eq. 1 and Eq. 8, the dynamics formulation of the thrusting system can be written as: (9) Where T y zx \u03b8 \u03b8 =  q is the general coordinates of the system. The thrusting system is commonly composed of multiple hydraulic cylinders, which can be regarded as a redundant parallel mechanism according to its working principle. As Fig. 2.2 shows: The mapping features between the load in each chain and the load on cutterhead is given by[6]: (10) 1 0 sin 0 cos 0 1 0 0 0 0 cos sin y y y z p y y z y y z \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8    \u22c5 \u22c5      = +           \u2212 \u22c5 \u22c5     \u03c9       ( )2,3 2,3 2,3 16 2,3 1 1 0 sin 0 cos 0 1 0 0 0 0 cos sin y y y z y p p p y z y y z i i e d p i \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 =    \u22c5 \u22c5       + + \u00d7          \u2212 \u22c5 \u22c5      = + \u00d7     e e e \u03c9 I \u03c9 e M F r       ( ) ( )\u02c6 \u02c6 \u02c6 0y z x \u03b8 \u03b8     + \u2212 \u2212 =     I K q M q F     e   \u22c5 = =     F G f F M 694 C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure1.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure1.4-1.png",
        "caption": "Fig. 1.4 a Influence of loads to each other. b Influence of voltages sources to each other",
        "texts": [
            "1 Volt\u2013Ampere Characteristics of an Active Two-Pole The simplest circuit of an active two-pole is shown in Fig. 1.3a. At change of a load resistance RL from the short circuit SC to open circuit OC, a load straight line or volt\u2013ampere characteristic in Fig. 1.3b is given by a linear expression IL \u00bc V0 Ri VL Ri \u00bc ISCL VL Ri ; \u00f01:1\u00de where Ri is an internal resistance and ISCL is the SC current. (a) (b)Fig. 1.3 a Simplest circuit. b Load straight line 1.1 Typical Structure and Equivalent Circuits \u2026 3 In turn, an internal resistance Ri in Fig. 1.4a determines the influence of the load resistances RL1; RL2 to each other. Similarly, the influence between the paralleling voltage sources V01; V02 takes place in Fig. 1.4b. Let us consider the straight lines of the initial circuit and a similar circuit with the other values ~V0, ~Ri in Fig. 1.5. The regimes of these circuits will be similar or equivalent if the correspondence of the characteristic and running regime parameters takes place. For the given case, this conformity is specified by arrows. 1.2.2 Regime Parameters in the Relative Form Let us constitute the relative expression of the load straight line for our simplest circuit in Fig. 1.3. For that, it is possible to use the values ISCL ; V0 as the scales"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001174_0954405414554016-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001174_0954405414554016-Figure8-1.png",
        "caption": "Figure 8. Dynamic simulation modelling of an angle-contact ball bearing: (a) initial status, (b) axis preloading, (c) radial preloading, and (d) axis and radial preloading.",
        "texts": [
            " In the simulation process, a non-flexible and non-massive auxiliary outer bearing is added and revolves at the same angular velocity as the inner ring even though the outer ring is actually static. The rolling bodies are regarded as springs, whose stiffness equals the composition stiffness of the rolling body in contact with the inner and outer rings (equation (31)), and these springs connect the auxiliary bearing\u2019s outer and inner rings. The spring\u2019s stiffness and deflection angle a can be adjusted according to the axial and radial preload forces (Figure 8). In radial preload, there is a derived axial force due to the existence of the contact angle, so that the angle-contact ball bearing has a derived displacement of axial. In axial and radial preloading, the axial displacement should include the axial displacement produced by itself and the derived one from radial preload which could be ignored. The dynamic simulation modelling actually facilitates the theoretical calculation formula of the axial and radial stiffnesses ka = kcA 0:5 XNb j=1 z 1\u00f0 \u00de0:5x2 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003114_icelmach.2016.7732535-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003114_icelmach.2016.7732535-Figure2-1.png",
        "caption": "Fig. 2. Relative position of the different modules with respect to one another. For sake of clarity, reference is made to a radial flux machine, but the same holds for axial flux machines. The indicated angles are electrical angles; moreover, they are measured with respect to the axis of phase 1-1. \u03b6 is the angle between the d-axis of the first rotor and the axis . The rotors show a reciprocal shift angle \u03c8, while the stators are aligned.",
        "texts": [
            " IV, a comparison between analytical and simulation results is performed. T The analysed multi-modular machine is represented in Fig.1. Several AFPM modules are keyed on the same shaft and each of them is connected to a two-level AC-DC converter. The converters are paralleled on the DC side. Two configurations are possible: all the stators aligned and the rotors phase shifted among them by an angle \u03c8; all the rotors aligned and the stators reciprocally phase shifted by \u03c8. The first condition will be analysed here (Fig. 2). The associated reference directions are adopted, even if the electrical machine operates as a generator. The assumed rated data of each module are reported in Table I. As can be understood on the basis of the rated speed, a direct-drive turbine-generator system is considered. Each module has a reasonably limited rated power, which is compatible with the use of two-level converters. The machine is isotropic and has neither damping cage nor significant damping equivalent effect; thus d and q axis parameters have the same values and the sub-transient inductance coincides with the synchronous one"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002539_1.g001799-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002539_1.g001799-Figure1-1.png",
        "caption": "Fig. 1 Seeker-head schematic diagram.",
        "texts": [
            " The study includes a development of the seekerhead dynamics model, conventional proportional-integral-derivative (PID), SMC, and supertwisting SMC (STSMC) controllers. The seeker LOS performance with the developed controllers was evaluated using numerical simulation and laboratory tests. The comparison clearly demonstrated the advantage of the nonlinear higher-order SMC methods in meeting system performance characteristics, given system uncertainty and disturbances. The low-cost seeker head considered in the current study is depicted schematically in Fig. 1. The inner and outer gimbals perform yaw (Y) and pitch (P) motion, respectively. The resulting attitude Euler angles are denoted by \u03c8 and \u03b8, respectively. Two microelectromechanical system (MEMS) gyros are mounted on the inner gimbal to measure the inertial angular velocities normal to the LOS. An approximate nonlinear dynamic model of the seeker head can be expressed as [17] _x f x g x u w x;\u03c9B; _\u03c9B d x (1a) y h x v x;\u03c9B (1b) Its four states are the attitude and angular rates of the two gimbals (i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002933_gt2016-56900-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002933_gt2016-56900-Figure4-1.png",
        "caption": "Figure 4. The basic structure of silicon iron sheet",
        "texts": [
            " Back-to-back centrifugal compressor (a) without HPS (b) with HPS Figure 3. The stator of electromagnetic actuator This new electromagnetic actuator needs a hole pattern on the inner surface of poles, and these holes lead to the reduction of the magnetic flow area and have negative effects on the mechanical properties of the actuator. ANSYS was applied to do the stable and transient electromagnetic field analysis for the AMP-HPS actuator to get the quantitative analysis for the effect of holes on the mechanical properties and eddy-current loss. Figure 4 shows the basic structure of silicon iron sheets that consist the AMB stator. Its dimensional parameters are listed in Tab.1. The rotor and stator of the acturator are laminated to reduce the eddy current losses in the integrated AMB-HPS. The Silicon iron is selected as the lamination material for both the rotor and the stator for its low magnetic losses and relatively low cost compared to those of other magnetic materials. The type of material used in the lamination of the rotor and stator is 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure7.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure7.10-1.png",
        "caption": "Fig. 7.10 (a) Procedures of fused deposition modeling and (b) a tabletop FDM machine. (Ultimaker 2+, reproduced with permission)",
        "texts": [
            " After the hardening process, the \u201cfresher\u201d overhead layer solidifies and bonds to the lower existing layer. Then, the layer fabrication repeats until the entire product is generated. Afterward, the product is taken out of the tank and rinsed with resin solvent in order to remove excess resin (after some time, the resining solution can also remove the support materials). Finally, the product is baked in an ultraviolet oven to ensure thorough curing of the photopolymer. Fused deposition molding (FDM) is a kind of 3D printing technology developed after the SL process, as shown in Fig. 7.10. The FDM technology was invented by Scott Crump in 1988 who founded his own company, Stratasys, based on the technology. In 1992, Stratasys introduced the world\u2019s first FDM-based 3D printer which also marked the transition of FDM technology into the commercial phase. Unlike other rapid prototyping technologies, FDM machines can be installed in the office environment which can significantly increase design and production efficiency in many industries. Fused deposition heats filamentous hot melt material and squeezes the material through an extruder with a fine nozzle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.42-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.42-1.png",
        "caption": "FIGURE 6.42",
        "texts": [
            " For reasons of habit, East is declared as zero, which corresponds to the X-axis in many schemes: \u00f06:23\u00de Note that the 2-argument arctangent function ATAN2 may be preferred over the simpler formulation since the trajectory can go in any direction. For trajectories that describe a loop, some care is needed as the calculations wrap through 180 . Distance travelled is simply formulated as the integral of the velocity vector. Noting that the orientation of the vector varies with time, the actual path on the ground can be calculated as X and Y: \u00f06:24\u00de \u00f06:25\u00de \u00f06:26\u00de For a vehicle travelling with an actual compass heading angle J at an actual distance travelled of S, from Figure 6.42, presuming small angle theory applies: \u00f06:27\u00de \u00f06:28\u00de \u00f06:29\u00de \u00f06:30\u00de Rearranging to find dS: Intrinsic coordinate calculation proposed. Reproduced from Odhams (2006). \u00f06:34\u00de \u00f06:35\u00de \u00f06:36\u00de In the limit, as dt tends to zero: \u00f06:37\u00de Similarly, it can be written by inspection that \u00f06:38\u00de Using the previous relationships: \u00f06:39\u00de \u00f06:40\u00de \u00f06:41\u00de \u00f06:42\u00de In the limit, as dt tends to zero: \u00f06:43\u00de In principle, the values for Jt, the derivative with respect to distance along the trajectory must be determined from the model and the differential equations integrated to give current values for actual distance travelled and path error"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001816_978-3-319-10401-0_22-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001816_978-3-319-10401-0_22-Figure1-1.png",
        "caption": "Fig. 1. Coordinate frames of manipulator",
        "texts": [
            " In other words, the angles obtained for each joint are used to compute the Cartesian coordinate for end effector. The training data of neural network have been selected very precisely. Especially, unlearned data in each neural network have been chosen, and used to obtain the training set of the last neural network. The Denavit-Hartenberg (D-H) notation and methodology are used in this section to derive the kinematics of robot manipulator. The coordinate frame assignment and the DH parameters are depicted in Fig. 1, and listed in Table 1 respectively, Xu et al., 2005. ))(()( ' txft =\u03b8 ),(2tan 331 xxyy adpadpa \u2212\u2212=\u03b8 2m + )( cos - )B ,atan2(B = 2 2 2 1 232 212 \u03c0\u03b8 BB acd a + + Where, 23222321 c )s(d + s )a + c(d = B 23222322 s )s(d - c )a + c(d = B and m = -1, 0 or1         \u2212\u2212+ 22 2 2 2 2 22 3 2 acos \u00b1= da dara z\u03b8 122232 dsasdrz +\u2212\u2212=       \u2212 \u2212=\u03b8 1 23z231z 23 z 4 s )c/osco( , c o 2tana ( ) ( ){ }zzzz asccnnsccaa 23423234235 ,2tan \u2212+\u2212=\u03b8 (3) (4) (5) (6) (7) Where(X, Y, and Z) represents the position and{ })a,a,a(),o,o,o(),n,n,n( zyxzyxzyx the orientation of the end-effector"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.45-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.45-1.png",
        "caption": "FIGURE 8.45",
        "texts": [
            "118d) These equations could be implemented using, for example, a Microsoft Excel spreadsheet for numerical solutions. The Y-position of the block, from 0 to 3.05 seconds is shown in Figure 8.44(b). This agreed very well with the SolidWorks Motion result in Figure 8.44(a). At time 3.05 sec, the Y-position of the block was 4.49 in., which matched well with that of SolidWorks Motion. The second example is the kinematic analysis of a single-piston engine. The engine consisted of four major components: case, propeller, connecting rod, and piston, as shown in Figure 8.45(a). The propeller was driven by a rotary motor at the angular speed of 60 rpm (i.e., one revolution per second). No gravity was present and the English units system was assumed. The engine was properly assembled with one free degree of freedom. When the propeller was driven by the rotary motor, it rotated, the crank shaft drove the connecting rod, and the connecting rod pushed the piston up and downwithin the piston sleeve. The engine assembly consisted of three subassemblies (engine case, propeller, and connecting rod) and one part (piston). The engine case was fixed (ground body). The propeller was assembled to the engine case using concentric and coincident mates, as shown in Figure 8.45(b). It was free to rotate along the X-direction. The connecting rod was assembled to the propeller (at the crankshaft) using concentric and coincident mates. It was free to rotate relative to the propeller (at the crankshaft) along the X-direction. Finally, the piston was assembled to the connecting rod (through the piston pin) using a concentric mate. The piston was also assembled to the engine case using another concentric mate. This mate restricted the piston\u2019s movement along the Y-direction, which in turn restricted the top end of the connecting rod to move vertically"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002683_978-3-319-11930-4-Figure6.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002683_978-3-319-11930-4-Figure6.6-1.png",
        "caption": "Fig. 6.6 Cross section of cup configuration with relative location of the ion plated Ni\u2013Cu\u2013Agcoated balls",
        "texts": [
            " The argon gas gets trapped inside the coating layers during deposition for process pressure of 17.5 mTorr and is liberated from the coating during the run-in portion of the RCF test. The rolling contact fatigue test platform presented in Chap. 4 was used to quantify the RCF life of the extreme DoE coated balls. The tests were conducted in high vacuum in the range of 10 7 Torr using a fixed load and speed for all tests. The Hertz contact stress was calculated as 4.1 GPa with a rotation speed of 130 Hz (7,800 rpm). For clarity, Fig. 6.6 contains a cross section of the cup and ball configuration and the proximity of the balls and rod for the RCF setup. This setup is the same as Fig. 4.1 in Chap. 4, except that the balls are coated with ion-plated nickel\u2013copper\u2013silver instead of evaporated pure silver. In addition to the validation tests of Chap. 4, the RCF test platform was further validated for repeatability using multielement coatings on 7.94 mm balls for nine tests. These tests used a separate set of nickel\u2013copper\u2013silver-coated balls, all from the same lot and process history"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001724_ssd.2015.7348203-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001724_ssd.2015.7348203-Figure6-1.png",
        "caption": "Fig. 6. Beam parameters and obstacle distance calculation",
        "texts": [
            "head(\u03c9) (5) This function takes into consideration the distance to obstacles \u201ddist (v,\u03c9)\u201d (giving preference of traveling longer distances without colliding with obstacles), the speed \u201dspeed(v)\u201d (preferring to travel at faster speeds) and heading \u201d head(\u03c9)\u201d (giving the progression of the system towards the goal). The \u03b1i values indicate the relative weight to be given to each term in the objective function. BCM obtains a divergent radial projection model of the environment based on the sensors\u2019 common position. The projection model is defined by Fig. 5. We can deduce ,from Fig.6, the geometric relations to calculate the angles: d = \u221a x 2 obs + (yobs) 2 \u2212 robs \u03b8obs = arctan ( xobs yobs ) \u03b8 = arcsin ( robs dobs ) \u03c11 = \u03b8obs \u2212 \u03b8 \u03c12 = \u03b8obs + \u03b8 The model is simplified and then a set of possible candidate beams is determined. After that, the best beam is calculated by maximizing the following objective function : f(\u03c11, \u03c12, d) = \u03b1 ( d cos(|\u03b5|) dmax ) \u2212 \u03b2 \u2223\u2223\u2223 \u03b5 \u03c0 \u2223\u2223\u2223 (6) where \u03b1 and \u03b2 are weight constants adjusted by experimentation with \u03b1+\u03b2 = 1 ; d cos(|\u03b5|)/dmax is the projected distance over the goal direction where \u03b5 is the angle between the goal direction and the robot\u2019s current orientation, defined as \u03b5 = \u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 0 if \u03c11 \u2264 \u03c10 \u2264 \u03c12 \u03c10 \u2212 \u03c12 if \u03c12 < \u03c10 \u03c11 \u2212 \u03c10 otherwise (7) where \u03c10 is the goal direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000804_1.4030097-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000804_1.4030097-Figure3-1.png",
        "caption": "Fig. 3 Crankshaft geometry",
        "texts": [],
        "surrounding_texts": [
            "SABRE-TEHL Simulation Tool. SABRE-TEHL is an inhouse advanced simulation tool which uses a finite difference 101510-2 / Vol. 137, OCTOBER 2015 Transactions of the ASME Downloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use numerical method coupled with an efficient Newton\u2013Raphson technique to solve the nonlinear Reynolds equation. In addition, the software includes both wear and asperity contact models based on the measured bearing surface roughness as well as full 3D solution of thermal energy equations and resultant temperature/ viscosity effects. The SABRE-TEHL takes into consideration a nonuniform clearance mainly due to \u2022 bearing design (including eccentricity and bore relief height) \u2022 bearing distortion due to assembly \u2022 static and dynamic misalignment of the crankshaft \u2022 crank journal machining operations such as barreling and lobbing [11]. SABRE-TEHL Input Data. In this study, two load speed cases have been considered. They relate to the engine running at \u201cmaximum torque\u201d (1447 rpm) and \u201cmaximum speed\u201d conditions (2045 rpm). Details of the engine input data together with the oil feed conditions are listed in Table 1. An additional in-house package has been used called SABREM. It is a \u201croutine\u201d bearing simulation tool for bearing design and material selection and used as a preliminary step for specialized simulations, i.e., SABRE-TEHL. This simulation tool uses the engine data to calculate loading for each bearing and predicts the bearing performance and temperatures for a complete set of crankshaft and connecting rod bearing. The bearing load as shown in Figs. 6 and 7 is been presented for both load speed conditions at different CA. Journal of Engineering for Gas Turbines and Power OCTOBER 2015, Vol. 137 / 101510-3 Downloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use In total, 45 thermo-EHL cases have been run. These included six different bearing geometries (Fig. 8) at both maximum torque and maximum speed conditions, two different eccentricities for the nongrooved standard bearing, increased oil groove widths and reduced clearances. Various oil groove depths have been simulated through the clearance defined groove method. It has to be noted that both nongrooved and grooved bearings were eccentrically bored with an estimated diametral clearance at crown of 109 lm. The bearing hole indicated in black is located at 345 deg bearing angle (BA) while the journal drilling hole (highlighted in ligher grey) is positioned at 42 deg BA. The lightest grey region denotes the circumferential oil groove extents."
        ]
    },
    {
        "image_filename": "designv11_64_0002448_ever.2016.7476406-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002448_ever.2016.7476406-Figure2-1.png",
        "caption": "Fig. 2. Cross-section of the example round-rotor machine.",
        "texts": [
            "A, we may want that B; E [0, 2n]. For this purpose, (34) can be written as per (36) withfdefmed by (37). B; = 8r(BY)) = mod( f (B;r) ) + 21f,21f) (36) f(BY\"l) = atan\ufffdKl + K2 CO\ufffdBd + BY) -aAsin(Bd + BY) -a)) (37) where atan2( a,b) is a function that returns an angle \u00a2 between -n and n such that tan( \u00a2)=b/ a. C. Relocation of slots in an example machine To clarify what has been discussed in the previous subsections, let us consider the example four-pole round rotor dual-three-phase synchronous machine shown in Fig. 2 and characterized by the data given in Table I. The same will be used to illustrate and assess (by FE analysis) the results presented in the next Sections as well. T ABLE I. EXAMPLE MACHINE CHARACTERISTIC DA T A Stator bore diameter, R, Rotor outer diameter, R, Core length, L Number of stator slots, Z, Number of rotor slots, Z,.(inc1uding \"dummy slots'\") Series-connected conductors per stator coil side, n, Series-connected conductors in rotor larger slots, nd Series-connected conductors in rotor smaller slots, n,2 215xlO-J m 208xlO-J m 750x10-J m 48 44 5 20 16 As illustrated in Fig",
            " The transformed geometry has a non-eccentric round rotor with non-uniformly distributed slots and is relatively simple to study through closed-form formulas. Results of its analysis can be then transferred to the original eccentric topology as discussed in the next three Sections which will respectively cover the computation of: inductances, air-gap field and UMP. The results of the analysis will be presented in a general form and will be illustrated and assessed by FE analysis by application to the example machine described in Fig. 2, Fig. 3 and Table I. In general, let us assume a machine equipped with N, stator windings WI(s) . . . WN/s) and Nr rotor windings WI(s) . . . WN/s) (each winding can be a whatever circuit like a phase, a coil, a parallel path of a phase or the rotor excitation field). Explicit close-form expressions will be derived in this Section for self and mutual inductances of machine windings. For this purpose, we shall use the fact that, according to conformal mapping theory, self and mutual inductances of circuits are invariant through conformal transformations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003099_iciea.2016.7603580-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003099_iciea.2016.7603580-Figure1-1.png",
        "caption": "Fig. 1 The coordinate system",
        "texts": [
            " The speed constraints are considered so that speed jump issues of backstepping control can be solved well. II. UUV Kinematic Model In order to study the UUV kinematic characteristics, a coordinate system of the UUV is established first. The coordinate system consists of two parts: inertial coordinate and vehicle coordinate. Inertial coordinate system E \u2212\u03be\u03b7\u03b6 is also called the Earth coordinate system or fixed coordinate system, and its origin is a certain point on earth. Vehicle coordinate system O xyz\u2212 is fixed on UUV and moves with UUV and it is also known as motion coordinate as shown in Fig. 1. Considering the convenience of calculation, vehicle coordinate system is introduced in the analysis of UUV\u2019s movement, since the moment of inertia and inertia of UUV are constant in vehicle coordinate, and the magnitude and direction of suffered thrust don\u2019t change with the changes of UUV\u2019s movement. The posture of UUV can be represented by position coordinates ( , , )x y z and attitude angle ( , , )\u03d5 \u03b8 \u03c8 in inertial coordinate; the linear and angular rates of UUV are represented by ( , , )u v w and ( , , )p q r respectively in the vehicle coordinate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002167_icelmach.2014.6960190-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002167_icelmach.2014.6960190-Figure4-1.png",
        "caption": "Fig. 4. Distribution magnetic flux density caused by pole changing.",
        "texts": [
            " The motor specifications used in the analytical model are given in Table \u2160. The variable magnetized magnet that we previously produced experimentally is a SmCo magnet [7]. The cost of the variable magnetized magnet appears to be the same as or lower than that of a conventional SmCo magnet because the variable magnet has the potential to decrease the volume of the required cobalt component. Magnetic field analysis confirmed that pole changing occurs in the rotor. The magnetic flux density distributions in the motor for both pole configurations are shown in Fig. 4. The magnetic flux forms an eight-pole distribution when all PMs are magnetized with the same polarity (Fig. 4a). In contrast, the magnetic flux forms a four-pole distribution when the polarities of the adjacent PMs oppose each other (Fig. 4b). The variable characteristics of the induced voltage are plotted in Fig. 5, and Fig. 6 shows the harmonic components of the induced voltage. The magnitude of the fundamental components will change from 60 to 100% when the motor flux changes from eight to four poles. The rotor can change its pole number when PM is magnetized. Magnetic field analysis was used to determine whether magnetization of the rotor PM could change and reverse the polarity. The magnetization was analyzed after the magnetizing current had flowed through the armature winding"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000812_2014-01-2256-Figure16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000812_2014-01-2256-Figure16-1.png",
        "caption": "Figure 16. Novel joint-based behavior acquisition system.",
        "texts": [
            " These effects are modelled using \u201cevent-based\u201d simulation, where the whole path is computed as preceding movement is a key input to future behavior (figure 14). The input forces into the machining process can be predicted based on real testing. Traditionally this is aquired using numerically controlled machine tools, and force-feedback units (figure 15). An innovative new method of obtaining the joint-based behavior of the robot was developed by Lund University. A patent has been successfully filed, and a spin-off is now commercializing the product. The set-up is shown in figure 16. The results from the backlash / input forces are then \u201cpluggedin\u201d to the CAM system to re-compute the robot program paths offline. The Delcam flowchart is shown in figure 17, with the recomputed robot path shown in figure 18. Applying both Kinematic-model and Joint-based-model coupled with good machining strategies at BTU, Germany did yield improvements compared to nominal - with results within 0.2mm (figure 19). As ever, the innovation behind the testing and the calculus took a significant time to develop"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003666_s00022-015-0274-2-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003666_s00022-015-0274-2-Figure2-1.png",
        "caption": "Figure 2 How to fold a (folded) rhombus R, a a (folded) rhombus, b an example of a state in the continuous folding process and c a flat folded state of R which is obtained by moving creases bq and cq where q moves from h to g1",
        "texts": [
            " We say that a family of polyhedra {Pt : 0 \u2264 t \u2264 1} is a continuous folding process from P = P0 to P1 if it satisfies the following conditions: (1) for each 0 \u2264 t \u2264 1, there is an intrinsic isometry from Pt onto P , (2) the mapping [0, 1] t \u2212\u2192 Pt \u2208 {Pt : 0 \u2264 t \u2264 1} is continuous, (3) P0 = P and P1 is a flat folded state of P . Let P = abcd be a regular tetrahedron. Then the two faces with common edge bc form a rhombus that is folded along bc; we denote this rhombus by R. Let g1 and g2 be centroids of those two faces. Apply Lemma 2.1 to the rhombus g1bg2c so that the moving creases are the line segments bq and cq where q moves from the midpoint h of bc to g2 (see Fig. 2). Lemma 2.3. [7] Let P = abcd be the regular tetrahedron in R 3 with vertices O = (0, 0, 0), a = (2/ \u221a 3, 0,\u22122 \u221a 2/ \u221a 3), b = ( \u221a 3, 1, 0), and c = ( \u221a 3,\u22121, 0). P is flattened explicitly by a continuous folding process of polyhedra {Pt : 0 \u2264 t \u2264 1} that satisfies the following: (i) the line segment Oh, where h is the midpoint of bc, is fixed on the x-axis, (ii) two faces abd and acd have no crease, (iii) there are points qt \u2208 Oh (0 \u2264 t \u2264 1) and rt \u2208 ah such that for each t, the face abc is divided into four triangles abrt, acrt, brth and crth, and bt(rt)\u2032h and ct(rt)\u2032h are attached to btqth and btqth, respectively, at, bt, ct and (rt)\u2032 denote the points on Pt corresponding to the points a, b, c and rt, respectively, (iv) the coordinates of at, bt, ct and qt are given by at = (6+2s \u221a 6+3s2\u221a 3(3+s2) , 0, 2(s\u2212\u221a 6+3s2) 3+s2 ), bt = ( \u221a 3, \u221a 1\u2212s2, s), ct = ( \u221a 3, \u2212\u221a 1\u2212s2, s), qt = ( \u221a 3(3+s2) 3+s \u221a 6+3s2 , 0, 0) where s = sin \u03c0 2 t (0 \u2264 t \u2264 1) (see Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000868_s12283-014-0166-y-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000868_s12283-014-0166-y-Figure3-1.png",
        "caption": "Fig. 3 Definitions of Euler angles that are used to describe the instantaneous attitude with respect to the inertial coordinate system",
        "texts": [
            "03, except around the stalling angle, while those for CL and CM are less than 0.05 and 0.03, respectively. The irregularity (standard deviation) becomes large near the stall angle. The inertial coordinate system is shown in Fig. 2. The origin is at the center of the throwing circle, with the XEaxis in the horizontal forward direction, the YE-axis in the horizontal lateral direction and the ZE-axis vertically downward. The coordinate system in the discus body-fixed system is denoted by xb, yb, and zb (Fig. 3c). The origin is defined as the center of gravity of the discus. It is assumed that the geometric center of the discus coincides with the center of gravity, that its zb-axis is aligned with the transverse axis (axis of symmetry), and that xb and yb are aligned with the longitudinal axes in the discus planform. Assuming the origin of the inertial coordinate system (XE, YE, ZE) is displaced without any rotation to the center of gravity of the discus, the new reference frame is defined as (X0, Y0, Z0) in Fig. 3a. The sequence of rotations conventionally used to describe the instantaneous attitude with respect to an inertial coordinate system is shown in Fig. 3 [7]. Starting with (X0, Y0, Z0) the following sequence is followed: (1) Rotate about the Z0-axis, nose right (positive yaw W, Fig. 3a); (2) Rotate about the y1-axis, nose up (positive pitch H, Fig. 3b); and (3) Rotate about the xbaxis, right wing down (positive roll U, Fig. 3c). Since there is a mathematical singularity (Gimbal lock) at H = 90 , quaternion parameters (q0, q1, q2, q3) [7] should be used instead of Euler angles when the flight trajectory is simulated. Therefore, the initial set of Euler angles is first transformed into the initial quaternion parameters [7]. We then have q0 \u00bc cos U=2\u00f0 \u00de cos H=2\u00f0 \u00de cos W=2\u00f0 \u00de \u00fe sin U=2\u00f0 \u00de sin H=2\u00f0 \u00de sin W=2\u00f0 \u00de \u00f03\u00de q1 \u00bc sin U=2\u00f0 \u00de cos H=2\u00f0 \u00de cos W=2\u00f0 \u00de cos U=2\u00f0 \u00de sin H=2\u00f0 \u00de sin W=2\u00f0 \u00de \u00f04\u00de q2 \u00bc cos U=2\u00f0 \u00de sin H=2\u00f0 \u00de cos W=2\u00f0 \u00de \u00fe sin U=2\u00f0 \u00de cos H=2\u00f0 \u00de sin W=2\u00f0 \u00de \u00f05\u00de q3 \u00bc cos U=2\u00f0 \u00de cos H=2\u00f0 \u00de sin W=2\u00f0 \u00de sin U=2\u00f0 \u00de sin H=2\u00f0 \u00de cos W=2\u00f0 \u00de \u00f06\u00de The force and moment equations of motion in the discus body-fixed system [7] are _U \u00bc 1 md Xa \u00fe 2mdg q1q3 q0q2\u00f0 \u00de\u00bd QW \u00fe RV \u00f07\u00de _V \u00bc 1 md Ya \u00fe 2mdg q2q3 \u00fe q0q1\u00f0 \u00de\u00bd RU \u00fe PW \u00f08\u00de _W \u00bc 1 md Za \u00fe mdg q2 0 q2 1 q2 2 \u00fe q2 3 PV \u00fe QU \u00f09\u00de _P \u00bc La IL QR IT IL 1 \u00f010\u00de _Q \u00bc Ma IL RP 1 IT IL \u00f011\u00de _R \u00bc Na IT \u00f012\u00de Here (U, V, W), (P, Q, R), (Xa, Ya, Za) and (La, Ma, Na) are the (xb, yb, zb) components of the velocity vector, the angular velocity vector, and the aerodynamic forces and moments, respectively",
            " The flight path angle, which is the elevation angle of the velocity vector, is denoted by c, and the azimuth angle is denoted by v. The magnitude of the velocity vector is assumed to be 26 m/s, which is reported as the maximum value for women [12]. It is self-evident that the higher the velocity, the longer the flight distance will be if the other release conditions are the same. The initial Euler angles, denoted by W, H and U (#3, #4 and #5) are used to describe the instantaneous attitude with respect to the inertial coordinate system (Fig. 3). Due to symmetry, there is no distinguishable difference between the xb-axis and the yb-axis. For this reason, it is impossible for Euler angles to obtain a unique optimal solution. Variables #6, #7 and #8 define the angular velocity vector. The spin rate about the transverse axis is denoted by R. Variable #9, the moment of inertia about the transverse axis (Fig. 5), is denoted by IT. The manufacturers of discuses have made great efforts to increase IT. It seems they believe that the larger IT (Angular momentum increases with IT if spin rate is constant) gives better attitude stability of a spinning discus"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002561_mawe.201600333-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002561_mawe.201600333-Figure10-1.png",
        "caption": "Figure 10. Schematic illustration of an extra-large rapid prototype",
        "texts": [
            " To investigate the relative error of the linear trend equation, nine test parts with different volumes of the support materials were fabricated for evaluation, Figure 9. The results of the relative error for nine test parts are shown, Table 6. The relative error can be controlled within 5.18 %. This means that the equation of y = 333.19x + 840.93 is suitable for predicting the removing time according to the volume of the support materials. To study the flexibility of the support materials removal system developed, an extra-large rapid prototype was fabricated for evaluation. The schematic illustration of extra-large rapid prototypes, Figure 10. The support materials of this rapid prototype can not be removed by ultrasonic machine due to limitation of container size. The extra-large rapid prototype with support materials is given, Figure 11. The volumes of the modeling materials and the sup- Figure 8. Removing time as a function of volume of the support materials Figure 9. Nine test parts for evaluating the relative error of removing the support materials from rapid prototypes using linear trend equation. White part and black part shown in the figure indicate the modeling materials and the support materials, respectively C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000887_gtindia2014-8186-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000887_gtindia2014-8186-Figure4-1.png",
        "caption": "Figure 4. Instrumentation system of SFD.",
        "texts": [
            "asmedigitalcollection.asme.org/ on 02/05/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 4 Copyright \u00a9 2014 by ASME retrieving such data, according to the prescribed formats. A manual mode of operation is also included with independent speed and load regulating potentiometers and functionally programmed push buttons. The proximity sensors give speed pulses based on the speeds of SGB shaft sides. Several parameters are measured using different instrumentations. SFD instrumentation system is shown in Figure 4. Rotor speed is measured through eddy current probes. Axial static load measurement is computed from fluid pressure. Shaft orbital movement along vertical and horizontal axis is monitored using eddy current probes. Film thickness is measured by eddy current probes. Vibration level is measured through accelerometers and charge amplifiers. Strain level is monitored using the strain gauge on SFD bearing webs. The arrangement of the test rig rotor is supported at two locations over a specified span length is shown in Figure 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure9-1.png",
        "caption": "Figure 9. Velocity-based detection mechanism.",
        "texts": [
            " Each Claw C has Guide Bar C. Four Guide Bars C are respectively inserted into four Guide Holes C of Plate C. Gear D meshes with Gear A. Rotary Damper is connected to Gear D. Claw D is connected to the shaft of Rotary Damper. One end of Linear Spring D is connected to Claw D, and another end is connected to Frame D. Frame D is mounted on Frame C. Switch C which can interrupt electric power supply to all motors of the robot is installed at the position of being pressed by Pin C when Plate C is rotated. Fig. 9 shows the mechanism which mechanically detects the unexpected robot motion on the basis of the angular velocity of Shaft C. The damping torque by Rotary Damper and the spring torque by Linear Spring D act on Claw D, when Gear D is rotated by Gear A. As the velocity of Gear A (i.e. Shaft C) increases, the damping torque increases. Claw D rotates by the torque difference between the damping torque and the spring torque, and locks Plate A, if the velocity of Shaft C exceeds the detection velocity level. The detection velocity level is adjustable by changing the attachment position of Linear Spring D as shown in Fig. 9. We can change the attachment position of Linear Spring D by using an adjustment mechanism of the detection velocity level as shown in [13]. Fig. 10 shows the mechanism to mechanically lock Shaft C. After Plate A is locked, each Claw B slides along each Guide Hole A of Plate A by the rotation of Plate B (Fig. 10(b)) and then one of three Claws B is hooked to the inner teeth and rotates Plate C (Fig. 10(c)). By the rotation of Plate C, Pin C switches off and each Claw C moves along each Guide Hole C (Fig",
            " 16 and Table II, we consider that the velocity-based safety device switched off the motor and locked Shaft C after the velocity of Shaft C approximately became the detection velocity level. The differences between the detection velocity levels and the experimental values are attributed to the damping torque errors of Rotary Damper, the attachment position errors of Linear Spring D, etc. Furthermore, in Fig. 16, the difference between the time when Claw D locked Plate A and the time when Shaft C was locked is attributed to the presence of the process of \u201cPlate A\u2019s stop (see Fig. 9 and Fig. 10(a))\u201d, \u201cClaw B\u2019s rotation (Fig. 10(b))\u201d , \u201cPlate C\u2019s rotation by Claw B (Fig. 10(c))\u201d, \u201c\u2018switch-off by Pin C\u2019 and \u2018Shaft C\u2019s lock by Claws C\u2019 (Fig. 10(d))\u201d which is required in the shaft-lock mechanism of the velocity-based safety device. The above experimental results indicate that the torque-based safety device can switch off the robot\u2019s motor, if the torque-based safety device detects the unexpected high torque even when the unexpected high velocity does not occur in the drive-shaft"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001510_s0263574713001203-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001510_s0263574713001203-Figure1-1.png",
        "caption": "Fig. 1. A two-link robot.",
        "texts": [
            " Similarly, as in the proof for Theorem 1, It is verified that Id, x, J (q), q\u0307 and J\u0307 (q) are bounded and x, Wk(t) ak, \u02dc\u0307q \u2192 0 as t \u2192 \u221e. From Eq. (22), it can be concluded that q\u0308 is bounded. Therefore, x\u0308 = J (q)q\u0308 + J\u0307 (q)q\u0307 is also bounded. Hence, x\u0308 is bounded if x\u0308d is bounded. This implies that x\u0307 is uniformly continuous; also we know that x is bounded, then according to Barbalat\u2019s Lemma, x\u0307 \u2192 0 as t \u2192 \u221e. In this section, we present simulation results of the proposed controller applied on a two-link RLED robot as shown in Fig. 1. The proposed controller is compared with the controller presented in ref. [20]. The dynamic model for this robot can be described in the form (1), (2) with M(q) = [ a + bcos(q2) c + b 2 cos(q2) c + b 2 cos(q2) c ] (84) C(q, q\u0307) = [\u2212 b 2 sin(q2)q\u03072 \u2212 b 2 sin(q2)(q\u03071 + q\u03072) b 2 sin(q2)q\u03071 0 ] (85) where a = m2l 2 2 + (m1 + m2)l2 1, b = 2l1l2m2, c = m2l 2 2 . l1 and l2 denote the lengths of the first and second links, respectively. q1 and q2 denote joint angle of the first and second links, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003145_978-981-10-2875-5_79-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003145_978-981-10-2875-5_79-Figure3-1.png",
        "caption": "Fig. 3 Schematic diagram of 3RRR",
        "texts": [
            " The manipulator used here is the one proposed in literature [11]. It is designed as the macro manipulator in the macro-micro manipulates system. Main target of this manipulator is realizing rapid precision positioning at the accuracy of 10 lm with the load of micro positioning platform. Once the manipulator is loaded, its kinematic parameters need to be re-calibrated. In this section, kinematic model of this mechanism will be introduced firstly, and then is the calibration model. Parameter identification also will be simply discussed. Figure 3 is the schematic diagram of 3RRR. As shown in this figure, Ai, Bi, Ci represent the revolute joints. AiBi and BiCi are the active and passive links, respectively. Their corresponding designed length are ai and bi. Orientation angles of AiBi and BiCi in base frame are hi and ci. Origins of frame Rb and Rm are the centroids of A1A2A3 and C1C2C3, respectively. Designed value of OAi and O0Ci are di and ci, their corresponding orientation angles are ai and bi. (i represents the number of the chains, its value is set from 1 to 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002357_978-3-319-17067-1_24-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002357_978-3-319-17067-1_24-Figure3-1.png",
        "caption": "Fig. 3 The description and analysis of DQV: a arrangement, b steering model",
        "texts": [
            " The height of the stride H in the transfer phase is as high as possible, and the length L in the support phase is as long as possible to ensure the walking ability. Based on the functional requirements, the 1-DoF planar leg mechanism with the close-chain full-pivot feature is presented in Fig. 2. The linkages in Watt chain [15] are designed as ground link, crank, thigh link, and shank link, respectively. In Fig. 2a, the parameters are described as: ri (i = 1, 2, \u2026, 9), the length of linages; \u03b8i,j (i = 1, 2,\u2026,9; j = 1, 2, \u2026,9), the angle between the linkages i and j. The design parameters are listed in Table 1. The arrangement of the DQV is shown in Fig. 3a. The whole mechanism consists of two identical quadruped mechanisms on each side of the frame. In each quadruped mechanism, the cranks arranged in the front and rear legs on the same side have a zero radian phase difference, and the cranks arranged in the left and right legs have a 180\u00b0 phase difference. The front and the rear legs follow the feature of moving in the same plane. In order to simplify the motion control algorithm and reduce the number of motors further, the cranks of the left and right legs share the public driving shaft, and the front and rear shafts are driven by the motor through the chain transmission. Consequently, each quadruped robot is driven independently by a single motor. Following the arrangement described above, the two quadruped mechanisms are equivalent to the two sides pedrail system, and they achieve different speed by changing the driving forces. Figure 3b depicts the no-load full-vehicle steering model. O and R are the steering center and the steering radius respectively, and B represents the distance between the longitudinal symmetry planes of the two quadruped robots. OT is the projection of point O in the longitudinal symmetry plane of the vehicle, and VT represents the average speed of the legged-vehicle. Similarly, V1 and V2 are the speed of the two quadruped mechanisms. Therefore, the steering parameters can be obtained as: xT \u00bc v1 R 0:5B \u00bc v2 R\u00fe 0:5B \u00bc v2 v1 B \u00f01\u00de R \u00bc v2 0:5BxT xT \u00bc v1 \u00fe 0:5BxT xT \u00bc v1 \u00fe v2 2xT \u00bc B v1 \u00fe v2\u00f0 \u00de 2 v2 v1\u00f0 \u00de \u00f02\u00de Figure 4 displays a cycle of walking simulation by ADAMSTM, the motor speed is 30r/min"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002495_s1068798x16010159-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002495_s1068798x16010159-Figure2-1.png",
        "caption": "Fig. 2. Calculation of the threaded joint by the methods in [1] (a) and [4] (b).",
        "texts": [
            " Experiments show that it is displaced relative to the neutral (geometric) axis of the contact plane toward the compressed side of the joint [1\u20135]. The use of variable pliability in considering the operation of a joint under a tipping torque was pro posed in [1]. In this approach, the threaded joint is replaced by z bushes (where z is the number of screws) connected by a rigid membrane, which is inclined to the contact planes of the bracing flanges. It is assumed that the membrane touches the contact plane on the compression side, but not on the breakaway side (Fig. 2a). The displacement x0 of the skew axis is assumed equal to the distance from the joint\u2019s neutral axis to the most remote screws on the side of the joint compressed by this torque. Assessment of the displacement was based on the following assumption in [4]: that the whole flange sur face operates in the compressed section of the joint but only the part of the surface in the vicinity of the screws operates on the breakaway section (Fig. 2b). In Fig. 3a, we plot experimental data for the relative contact increment \u03b4 of the plane surfaces under the action of the force F normal to the contact surface for 40X steel samples [7]. Lathe finishing of the contact surfaces ensures surface roughness Ra = 1.25 \u03bcm of Keywords: threaded joint, tipping torque, skew axis, contact deformation DOI: 10.3103/S1068798X16010159 RUSSIAN ENGINEERING RESEARCH Vol. 36 No. 1 2016 POSITION OF THE SKEW AXIS OF A THREADED JOINT 17 the steel samples. In Fig. 3b, we show analogous curves of \u03b4 as a function of the pressure p for unquenched steel 45 samples [7]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001707_10255842.2015.1070592-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001707_10255842.2015.1070592-Figure1-1.png",
        "caption": "Figure 1.\u00a0schematic illustration of the position corresponding to the three additional weights conditions.",
        "texts": [
            " Before the experiment, the subjects warmed up by executing forehand strokes during 10\u00a0min. The experimental task consisted of performing forehand strokes on balls shot by a ball machine at a constant speed. Subjects were required to send 3 balls at a vertical target (45\u00a0cm\u00a0\u00d7\u00a045\u00a0cm) located 0.92\u00a0m above the ground about 7\u00a0m away. The task was performed using three different conditions of racket weight distribution. The racket weight distribution was adjusted by adding 15\u00a0g at the end cap, at the throat and at the top cap (Figure 1). External forces applied by the hand were computed using a pressure map sensor (Hoof 3200, TekScan, Boston, Massachusetts, USA) wrapped around the handle with a 500-Hz sample acquisition. To evaluate the amount of grip force used in each stroke, the maximum grip force (MGF) and the impulse of grip force (IGF) were identified. The IGF was calculated as follows: (1)IGF = t 2 \u222b t 1 F \u22c5 \u0394t \u00a9 2015 taylor & Francis CONTACT J. rossi mail.jeremy.rossi@gmail.com D ow nl oa de d by [ U ni ve rs ity o f O ta go ] at 1 2: 59 1 7 O ct ob er 2 01 5 ComputER mEthods in BiomEChaniCs and BiomEdiCaL EnginEERing 2045 indicated that the end-cap and top-cap conditions were significantly higher compared to the throat condition (p\u00a0<\u00a00"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002189_phm.2014.6988158-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002189_phm.2014.6988158-Figure2-1.png",
        "caption": "Fig. 2. Planetary gearbox experimental system",
        "texts": [
            " 0 1 0 2 y y X y y X \u03c9 \u03c9 > \u21d2 \u2208\u23a7 \u23a8 < \u21d2 \u2208\u23a9 (13) For the purpose of validate the effectiveness of the method this paper proposed, an introduced fault experiment of planetary gearbox was implemented rely on the RCM laboratory of Mechanical Engineering College. The experimental system consists of a planetary gearbox, adrive motor, a speed and torque sensor which used to acquire the speed and torque information and a magnetic powder brake to provide load. And between the two of them is connected by coupling, as shown in Fig. 2. The wear fault is implemented on one tooth of sun gear. Fig. 3 shows the sun gear after machine the fault and its original photo. Four accelerometers are mounted on the planetary gearbox casing by glue, wherein accelerometer 1 and 2 are mounted on the input side of the gearbox (1 is horizontal and 2 is vertical), accelerometer 3 is on the top of the casing and accelerometer 4 is fixed on the output side. The specific location of every accelerometer is as depicted in Fig. 4, and Fig. 5 is the structure of the planetary gearbox"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003276_physreve.94.063002-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003276_physreve.94.063002-Figure2-1.png",
        "caption": "FIG. 2. Partial inversion of four different cones using finite elements. The top row shows full cones partially inverted. The middle row shows cutaway counterparts highlighting the crease width around the ridge line according to color changes in the maximum principal strain contours. The bottom row shows axisymmetrical meridians (green) compared to the perfectly straight meridians (gray), showing how the actual crease region diverges and detaches from the latter; the crease width is later defined to be the arc length of the detached region. The cone angles are (a) 60\u25e6, (b) 45\u25e6, (c) 30\u25e6, and (d) 0\u25e6; other material and geometric properties are the same as Fig. 1. In (d) the inverted part overlays the original undeformed tube, giving a mottled complexion of interference.",
        "texts": [
            "3 and a Young modulus of 1 MPa, which is relatively soft but intends to mimic a soft rubber, and all thicknesses are of the order of 0.1 mm for a cone side length of around 50 mm. The cone apex is replaced by a spherical cap of radius 3 mm, which connects smoothly to the rest of the cone; a vertical force is applied normal to the cap at its pole and the base of the cone is fully built in. Geometrical nonlinearity is coupled with a Riks arc length algorithm [7] during solution to capture fine, sometimes highly nonlinear features of the load path. Before examining a typical path, consider the four partially inverted cones in Fig. 2. They begin with a shallow cone in Fig. 2(a) inclined initially to the vertical by 60\u25e6 and become increasingly steeper with cone angles of 45\u25e6, 30\u25e6, and 0\u25e6, which is a cylindrical tube. Each panel also reveals the deformed cross section where the crease profile is clearly evident and accentuated by highlighting elastic strain contours; each inverted cap is also highly strained. A sense of mirror symmetry about the ridge plane is also conveyed, in particular for the tube, where the inverted part is passing through its undeformed self. We have also added the Pogorelov outline to each panel, in order to highlight the crease shape and its relative size as the cone angle decreases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.45-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.45-1.png",
        "caption": "FIGURE 6.45",
        "texts": [
            " Some care is needed in handling the wrapping of trigonometric function and with iteration convergence, as with all numerical approaches. It becomes inevitable with any form of vehicle dynamics modelling that the interaction of the operator with the vehicle is a source of both input and disturbance. In flight dynamics, the phenomenon of \u2018PIO\u2019 e pilot-induced oscillation e is widely known. This occurs when inexperienced pilots, working purely visually and suffering from some anxiety, find their inputs are somewhat excessive and causing the aircraft to, for example, pitch rhythmically instead of holding a constant altitude (Figure 6.45). PIO is caused when the operator is unable to recognise the effects of small control inputs and therefore increases those inputs, before realising they were excessive and reversing them through a similar process. It is analogous to the \u2018excess proportional control gain oscillation\u2019 discussed in classical control theory (Leva et al, 2002). For road vehicles, drivers most likely to induce PIO in steering tend to be inexperienced or anxious drivers travelling at a speed with which they are uncomfortable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001395_1.4029831-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001395_1.4029831-Figure1-1.png",
        "caption": "Fig. 1 (a) Experimental setup with powder feed system and (b) experimental setup with preplaced powder",
        "texts": [
            " The parameters, Contributed by the Manufacturing Engineering Division of ASME for publication in the JOURNAL OF MANUFACTURING SCIENCE AND ENGINEERING. Manuscript received June 27, 2014; final manuscript received February 13, 2015; published online March 2, 2015. Assoc. Editor: Z. J. Pei. Journal of Manufacturing Science and Engineering JUNE 2015, Vol. 137 / 031010-1 Copyright VC 2015 by ASME Downloaded From: http://manufacturingscience.asmedigitalcollection.asme.org/ on 03/01/2015 Terms of Use: http://asme.org/terms including defocus length and laser speed, were selected from a literature survey for good cladding quality [15]. Figure 1(a) shows the experimental setup for the first set of experiments with one gravity powder feed system. The clad track was formed by moving the specimen relative to the laser beam in the y-direction, while the powder flowed down from the feed system hopper. The second experiment shown in Fig. 1(b) entailed placing preplaced powder on the surface of the substrate material, and then the laser light hitting the preplaced powder on the substrate material surface. Three samples, B1 (10% SiC, 10 mm defocus length, 500 mm/min speed), B2 (15% SiC, 15 mm defocus length, 500 mm/min speed), and B3 (20% SiC, 15 mm defocus length, 800 mm/min speed), were used in this experiment as illustrated in Tables 3 and 4. 3.1 Experiment 1: One Powder Feed System Containing Different SiC% With P25 (Iron-Based Matrix Material)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001214_ijhvs.2014.068101-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001214_ijhvs.2014.068101-Figure10-1.png",
        "caption": "Figure 10 Flange\u2013rail contact points",
        "texts": [
            " Slip velocities at the contact \u2018a\u2019 are obtained as follows: Forward slip, 3 3a aS V= (36) It is assumed that there is no tangential displacement of rail. Lateral slip, 1 1 1 .a a t aS V V= \u2212 (37) Spin component total angular velocity of the wheelset, 2 \u02c62 .a aXYZ S S= \u2126 \u22c5 (38) Thus, the angular velocity component of the wheelset about the normal to the patch plane at contact \u2018a\u2019 can be realised in a bond graph from ,\u03b8 \u03c9 and \u03c8 using transformers of moduli obtained from equation (38). Flange is considered to be a circular disc of radius Rf with its centre at fao on the wheelset axis as shown in Figure 10. For zero roll angle (\u03b8), two possible flange-rail contact points a1 and a2 are the points of intersection of the horizontal plane passing through the contact point \u2018a\u2019 and the flange circle. Since conicity and flange to rail clearance are small, these two points are assumed to be possible contact points at other times also. To determine the flange-rail contact at, say, \u2018a1\u2019, radial component of the velocity of a1 is integrated from given initial condition and compared with flange-rail clearance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001258_j.wear.2015.01.039-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001258_j.wear.2015.01.039-Figure8-1.png",
        "caption": "Fig. 8. Steps of bristle/loop deformation during contact with the slider: (a) before deformation, (b) during bending deformation, and (c) during sliding underneath the slider. F is the total friction force, x is the slider displacement, h is the slider altitude, and \u03a8 is the angle between the axis of the slider displacement and the normal to the bristle at the contact point.",
        "texts": [
            " The influence of the measurement conditions, that is, the sliding velocity and slider altitude, and the properties of the pile surfaces on the friction behaviour is studied. The typical evolution of the tangential and normal forces relative to the displacement is presented in Fig. 6 for a single pile entity (single loop or single bristle) and in Fig. 7 for an assembly of several loops. The same curve shape is obtained from the nanotribometer and the macrotribometer. During contact with the slider, the pile entity is first bent, that is, for 0r\u03a8r\u03c0/2, and then rubbed under the slider, that is, \u03a8\u00bc\u03c0/2 (Fig. 8), where \u03a8 is the angle between the axis of the slider displacement and the normal to the bristle at the contact point [23]. Firstly, Fig. 6 illustrates the evolution of the normal and the tangential forces. This evolution can be divided into two parts: Zone no. 1 is characterized by a high peak, which mainly corresponds to the bending mechanism of the bristle/loop: it is the deformation part (0r\u03a8r\u03c0/2). Zone no. 2 is the plateau. It corresponds to the friction under the slider (\u03a8\u00bc\u03c0/2). The length of this plateau corresponds to the length of the slider",
            " The curves obtained are similar to those obtained by considering a single bristle with the nanotribometer. Two different criteria have been extracted from these curves: the maximum force of the total friction force, that is, the peak due to the bristle/loop bending zone and the COF between the bristle/loop and the slider. In accordance with the theory [24], the total friction force F (Fig. 9) increases when the vertical position h of the slider decreases. In fact, according to Watzky et al. [23], the total friction force is given by the following formula (Fig. 8): F \u00bc 2UBU sin \u03a8 hU cos \u03a8 \u00fexU sin \u03a8 2 \u00f01\u00de where \u03a8 is the angle between the axis of the slider displacement and the normal to the bristle at the contact point, B is the bending rigidity of the pile entity, h corresponds to the altitude of the slider, x the sliding distance. B, a material parameter, and h, an experimental adjustment, are constant during the experiment. \u03a8 and x vary with the slider displacement. The maximum of the total force is obtained for a low value of\u03a8 and a sliding distance d"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000640_icra.2015.7139415-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000640_icra.2015.7139415-Figure1-1.png",
        "caption": "Figure 1.VOOPS Configuration",
        "texts": [
            " However, this results in an increase in size (the size becomes 24% more) and the power consumption increases by 50% compared to quadrotor system. Therefore, increasing the number of rotors or size of rotors is not a good option here. Use of a larger propeller without increasing the footprint is not possible in conventional way of designing quadrotors, where the quadrotor tip to tip size cannot be smaller than (1+\u221a ) times the diameter of the propeller. To overcome this limitation, the new configuration of VOOPS is proposed, where propellers are placed in different planes allowing overlaps as shown in Fig. 1. In this way, it is possible to accommodate larger propellers in this design, resulting in better endurance and payload carrying capability for the same size. The quadrotor configuration discussed in this paper is an outcome of the Micro Aerial Vehicle Challenge 2014, India [4]. The mission demanded a payload capacity of 150 grams with a flight time of about 20 minutes. The system had to be accommodated in an infantry case of inner dimension 300 mm sphere. This led to a unique configuration VOOPS",
            " Though size is our prime constraint, smaller being better to enter through doors/windows, the smallest design might not be efficient compared to the bigger ones with less overlap. Choosing the optimal design is a challenge. In this paper we have designed the quadrotor such that it does not exceed the given size restriction. The quadrotor was designed with 22 mm overlap on all the propellers such that the dimension of the quadrotor is within the specified limit, which gives a percentage overlap of 15.7%. The experimental quadrotor designed with the above parameters is shown in Fig. 7. It is convenient to define two coordinate frames as indicated in Fig. 1 to analyse the 6-DoF motion of a VOOPS Quadrotor system [9]. The coordinate frame * + is the body-fixed reference frame and * + is the earth fixed or inertial frame. \u03b7 is defined as the pose vector of the system represented by 1 2[ , ]T T T\u03b7 \u03b7 \u03b7 where is the position vector given by =[ , , ]Tx y z 1 \u03b7 and is the orientation vector given by 2 =[ ]T   \u03b7 where     are roll, pitch and yaw respectively. v is the velocity vector represented by 1 2 [ , ]T T T v vv where 1v is the linear velocity vector given by 1 [ , , ]Tu v wv and 2v is the angular velocity vector given by 2 [ , , ]Tp q rv "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001891_j.ifacol.2015.12.111-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001891_j.ifacol.2015.12.111-Figure1-1.png",
        "caption": "Fig. 1. Inertial frame and body frame of the quadcopter",
        "texts": [
            " Section 3 presents the theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experimental data using the frequency-sampling filter (FSF). Section 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNAMICS The closed-loop dynamics of a quadcopter are introduced first. The reference systems used in this paper are shown in Fig. 1. The quadcopter\u2019s attitude is defined with three Euler angles, namely roll about x-axis, pitch about y-axis, and yaw about z-axis. For simplicity, in this paper the relay feedback experiment is performed on the rotation 17th IFAC Symposium on System Identification Beijing International Convention Center October 19-21, 2015. Beijing, China Copy ight \u00a9 IFAC 2015 122 Step Response Identification of a Quadcopter UAV Using Frequency-sampling Filters Xi Chen \u2217 Liuping Wang \u2217\u2217 \u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia \u2217\u2217 School of Electrical and Computer Engineering, RMIT University, M lbourne,Victoria 3000, Australia (e-mail: liuping",
            " Section 3 present theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experim al data using the frequency-sampling filter (FSF). S ction 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNAMICS The closed-loop dynamics of a quadcopter are introduced first. Th reference systems used in this paper are shown in Fig. 1. Th quadcop r\u2019 attitude is defined with three Euler angles, namely roll abou x-axis, pitch abou y-axis, and y w about z-axis. For simplicity, in this paper the relay feedback experiment i performed on the rotation 17th IFAC Symposium on System Identification Beijing Internati nal Convention Center October 19-21, 2015. Beijing, China Copyright IFAC 2015 122 Step Response Identification of a Quadcopter UAV Using Frequency-sampling Filters Xi Chen \u2217 Liuping Wang \u2217\u2217 \u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia \u2217\u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia (e-mail: liuping",
            " Section 3 presents the theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experimental data using the frequency-sampling filter (FSF). Section 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNAMICS The closed-loop dynamics of a quadcopter are introduced first. The reference systems used in this paper are shown in Fig. 1. The quadcopter\u2019s attitude is defined with three Euler angles, namely roll about x-axis, pitch about y-axis, and yaw about z-axis. For simplicity, in this paper the relay feedback experiment is performed on the rotation 17th IFAC Symposium on System Identification Beijing International Convention Center October 19-21, 2015. Beijing, China Copyright \u00a9 IFAC 2015 122 t s s I ti ti f t si -s li ilt s i hen \u2217 iuping ang \u2217\u2217 \u2217 School of lectrical and o puter ngineering, I niversity, elbourne, ictoria 3000, ustralia \u2217\u2217 School of lectrical and o puter ngineering, I niversity, elbourne, ictoria 3000, ustralia (e- ail: liuping",
            " Section 3 presents the theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experimental data using the frequency-sampling filter (FSF). Section 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNA ICS The closed-loop dynamics of a quadcopter are introduced first. The reference systems used in this paper are shown in Fig. 1. The quadcopter\u2019s attitude is defined with three Euler angles, namely roll about x-axis, pitch about y-axis, and yaw about z-axis. For simplicity, in this paper the relay feedback experiment is performed on the rotation 17th IFAC Symposium on System Identification Beijing International Convention Center October 19-21, 2015. Beijing, China Copyright \u00a9 IFAC 2015 122 Xi Chen et al. / IFAC-PapersOnLine 48-28 (2015) 122\u2013127 123 Step Response Identification of a Quadcopter UAV Using Frequency-sampling Filters Xi Chen \u2217 Liuping Wang \u2217\u2217 \u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia \u2217\u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia (e-mail: liuping",
            " Section 3 presents the theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experimental data using the frequency-sampling filter (FSF). Section 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNAMICS The closed-loop dynamics of a quadcopter are introduced first. The reference systems used in this paper are shown in Fig. 1. The quadcopter\u2019s attitude is defined with three Euler angles, namely roll about x-axis, pitch about y-axis, and yaw about z-axis. For simplicity, in this paper the relay feedback experiment is performed on the rotation 17th IFAC Symposium on System Identification Beijing International Convention Center October 19-21, 2015. Beijing, China Copyright \u00a9 IFAC 2015 122 Step Response Identification of a Quadcopter UAV Using Frequency-sampling Filters Xi Chen \u2217 Liuping Wang \u2217\u2217 \u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia \u2217\u2217 School of Electrical and Computer Engineering, RMIT University, M lbourne,Victoria 3000, Australia (e-mail: liuping",
            " Section 3 present theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experim al data using the frequency-sampling filter (FSF). S ction 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNAMICS The closed-loop dynamics of a quadcopter are introduced first. Th reference systems used in this paper are shown in Fig. 1. Th quadcop r\u2019 attitude is defined with three Euler angles, namely roll abou x-axis, pitch abou y-axis, and y w about z-axis. For simplicity, in this paper the relay feedback experiment i performed on the rotation 17th IFAC Symposium on System Identification Beijing Internati nal Convention Center October 19-21, 2015. Beijing, China Copyright \u00a9 IFAC 2015 122 Step Response Identification of a Quadcopter UAV Using Frequency-sampling Filters Xi Chen \u2217 Liuping Wang \u2217\u2217 \u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia \u2217\u2217 School of Electrical and Computer Engineering, RMIT University, Melbourne,Victoria 3000, Australia (e-mail: liuping",
            " Section 3 presents the theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experimental data using the frequency-sampling filter (FSF). Section 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNAMICS The closed-loop dynamics of a quadcopter are introduced first. The reference systems used in this paper are shown in Fig. 1. The quadcopter\u2019s attitude is defined with three Euler angles, namely roll about x-axis, pitch about y-axis, and yaw about z-axis. For simplicity, in this paper the relay feedback experiment is performed on the rotation 17th IFAC Symposium on System Identification Beijing International Convention Center October 19-21, 2015. Beijing, China Copyright \u00a9 IFAC 2015 122 t s s I ti ti f t si -s li ilt s i hen \u2217 iuping ang \u2217\u2217 \u2217 School of lectrical and o puter ngineering, I niversity, elbourne, ictoria 3000, ustralia \u2217\u2217 School of lectrical and o puter ngineering, I niversity, elbourne, ictoria 3000, ustralia (e- ail: liuping",
            " Section 3 presents the theory and implementation of quadcopter data acquisition using the relay feedback experiment. Section 4 describes the identification of the system step response from the experimental data using the frequency-sampling filter (FSF). Section 5 presents experimental results of the quadcopter step response identification. Section 6 concludes this paper. 2. QUADCOPTER CLOSED-LOOP DYNA ICS The closed-loop dynamics of a quadcopter are introduced first. The reference systems used in this paper are shown in Fig. 1. The quadcopter\u2019s attitude is defined with three Euler angles, namely roll about x-axis, pitch about y-axis, and yaw about z-axis. For simplicity, in this paper the relay feedback experiment is performed on the rotation 17th IFAC Symposium on System Identification Beijing International Convention Center October 19-21, 2015. Beijing, China Copyright \u00a9 IFAC 2015 122 about the x-axis only, and the same experimental set-up can be applied to the other two axes easily. The block diagram of the quadcopter closed-loop attitude control system is shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003145_978-981-10-2875-5_79-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003145_978-981-10-2875-5_79-Figure1-1.png",
        "caption": "Fig. 1 System setup",
        "texts": [
            " Kinematic calibration of 3RRR is introduced in Sect. 3. Section 4 is the experiment. Section 5 is the conclusion. The pose of the mobile platform can be measured by camera through recognizing low-level geometric features on it, such as the points or the lines [10]. In practical, point feature is more commonly used. As the 3RRR have three degree of freedoms in the motion plane, so theoretically two markers are enough. However, more marker points will make the measurement more robust to noise. A system schematic diagram is shown in Fig. 1. In this scheme, optical axis of the camera is set to be perpendicular to the base platform. Mobile platform can move to the desired pose under the control of the controller, and its actual pose can be measured through the camera and sent to the controller immediately. Hence, the automatic online calibration can be realized. In this section, we outline the main principles of the vision-aided online measuring system, the planar 3RRR manipulator used in this system is the one proposed in reference [11]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001630_j.proeng.2014.03.133-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001630_j.proeng.2014.03.133-Figure2-1.png",
        "caption": "Fig. 2. Deformation of flat TRME: a) deformation of packet, b) deformation of bond rubber layer.",
        "texts": [
            " For these structures in the low-deformation domain (up to 5% \u00f7 10%) high intensity of the external load (up to 200 MPa) may be exerted in practice. Experimental studies [7 \u00f7 14] indicate that under these loads a significant nonlinearity of the \"force-displacement\u00bb stiffness characteristics associated with the physical nonlinearity of elastomeric materials take place. Traditional methods of calculation [5] do not allow to describe non-linear stiffness characteristics of these elements. Deformation of TRME under axial force normal to flat surface is shown in Fig. 2. For determination of stiffness characteristics for laminated elastomeric structures with regard to the physical nonlinearity it is more suitable to use variational methods. It is assumed that the geometry of the elastomeric layer, which allows imposing a significant external load, provides small deformation, i.e. problem remains geometrically linear For boundary value problems of static theory of elasticity for low volume compressible materials only the physical group of equations - the ratio between stress tensor ij and strain tensor ij changes [5]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000812_2014-01-2256-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000812_2014-01-2256-Figure7-1.png",
        "caption": "Figure 7. Coordinate tracking (single LED)",
        "texts": [
            " The other options are simplified versions of joint excitation. A solid state camera system was used from Nikon Metrology. It measures strobed IR LED's at frequencies up to 1kHz with latency \u22483ms, using linear CCD's and sub-pixel interpolation. As part of its manufacture the camera has a calibration grid created. This is by having LEDs on a very large Coordinate Measuring Machine running through its full working volume. In operation if a LED is introduced into its field of view it measures it as a coordinate, as shown in Figure 7. With three or LEDS a 6D Frame is measured, as shown in figure 8. Multiple frames can be tracked simultaneously, allowing many measurement opportunities including relative motion tracking and nested tracking. This is ideal for robotics - as they are programmed with two coordinate reference systems (tool and base) To perform the kinematic modelling, the robot is moved through a series of random points and measured. The target vs actual points are compared in cartesian space (set-up shown in figure 9)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003244_s1068366616040073-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003244_s1068366616040073-Figure4-1.png",
        "caption": "Fig. 4. (a) Schematic of experiment on pressing of ridge into perfectly plastic solid moved perpendicularly to direction of pressing, (b) velocity hodograph for experiment on pressing of ridge, and (c) dependence of relative force of rolling PX/2K on parameter vX/vZ.",
        "texts": [
            " The real influence of the relative parameter vtool/vcont on the CR process is also confirmed by the fact that an increase in the velocity of rolling has little or no effect on the conditions for the stable occurrence of the CR process. In order to confirm the obtained regularity of the self-setting of the forces of friction during CR due to changes in the slippage of the billet against the tool, we consider and solve a problem of pressing a ridge into a perfectly plastic solid, which moves perpendicular to the direction of pressing. The simulation of the contact surface of a tool by an ensemble of sequentially arranged ridges is a known method used in studying the regularities of friction [10]. Figure 4a shows the schematics of the experiment. Ridge 4 with an apex angle of 90\u00b0 is pressed into perfectly plastic solid 1 with velocity vZ. The perfectly plastic solid moves perpendicular to the direction of the movement of the ridge with the velocity vX. Let us select the plastic area of the deformation zone ABCD, which consists of rigid areas 2 and 3. The problem is solved using the upper limit estimate method. Varying the positions of points A and D has made it possible to determine the optimum values of these positions, which ensures the minimum distortion-strain energy. Figure 4b shows the velocity hodograph for the problem. The solution is found for the force of rolling PX/2K, which in the case of CR is equivalent to the force of friction. Varying the value of the ratio vX/vZ from 0.5 to 3 yields values of the force PX/2K. The result is shown in Fig. 4c. The obtained theoretical solution confirms the experimental result; i.e., an increase in the slippage expressed in the relative parameters (vX/vZ) leads to an increase in the force of friction. As a result, we can consider it to have been proved that the self-setting of the forces of friction during CR is governed by the dimensionless parameter, i.e., the ratio of the velocity of the tool to the velocity of the flow of the metal in the near-contact layer of the deformation zone. Let us consider the relationship between the velocities near the contact zone and Prandtl\u2019s coefficient of friction \u03bc\u03c4, which was obtained using the slip-line technique (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure13-1.png",
        "caption": "Figure 13. Developed walking support robot.",
        "texts": [
            " When the magnitude of the Shaft A\u2019s torque does not exceed the detection torque level, the torque is transmitted to Shaft B via Hub, Balls, Center Flange and Parts A. If the magnitude of the torque exceeds the detection torque level, Balls pop out of V-shaped Pockets and Plate Z slides along Hub and switches off (Fig.12). The detection torque level is adjustable by changing the attachment position of Adjusting Nut. We developed the walking support robot equipped with four velocity-based mechanical safety devices and two torque-based mechanical safety devices. Fig. 13 shows the developed robot. As shown in Fig. 13, the length is 125[cm] and the width is 154[cm]. The armrest is adjustable in height from 85[cm] to 108[cm] according to the height of each patient by using a hand crank. We conducted the following experiments by using the developed walking support robot. We experimentally examined whether the torque-based safety device can achieve the function. Fig. 14 (a) shows the experimental setup. We mounted the walking support robot on the two spacers. Then, we attached a steel bar to the right wheel and brought the steel bar into contact with the block"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001949_s00397-015-0902-7-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001949_s00397-015-0902-7-Figure2-1.png",
        "caption": "Fig. 2 a Modified section of the UTM; b a schematic of the modified UTM and visualization setup",
        "texts": [
            "0 mN/m, at a ratio of 54:46 (wt%). By comparing the thinning behavior of these two fluids, the effect of the particles on a thinning filament under extensional flow could be investigated with nearly identical viscosity and surface tension, which govern the thinning dynamics of Newtonian fluids. To induce extensional flow, we modified a universal testing machine (UTM; LF plus, Lloyd Instruments, UK). We set up two circular plates with diameters of 3 mm on the top and bottom of the UTM as depicted in Fig. 2a. Extensional flow was induced by moving the upper plate upwards. The gap between the two plates was 1.5 mm and the moving speed of the upper plate was 1 mm/s. We used a high-speed camera (Photron Fastcam Ultima 512, Photron, USA) and a tenfold objective lens as illustrated in Fig. 2b to visualize the final stage of the filament breakup. The high-speed camera captures 4000 frames per second (fps) with a resolution of 256 \u00d7 512 pixels. One pixel of the recorded image corresponds to an area of 1.94 \u03bcm \u00d7 1.94 \u03bcm. An image having a resolution of 256 \u00d7 512 pixels represents an area of 486.4 \u03bcm \u00d7 972.8 \u03bcm. A 250-W halogen light (BMH-250, Mejiro precision, Japan) was used for backlight illumination. From this visualization setup, we could observe the thinning filament even in the length scale of a particle diameter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure5.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure5.9-1.png",
        "caption": "Fig. 5.9 Influence of the element regularity",
        "texts": [
            " Now we need to calculate the partial derivatives of the parametric (\u03be, \u03b7) coordinates with respect to the physical (x, y) coordinates, see Eqs. (5.71)\u2013(5.74). For case (a), this evaluation gives: 260 5 Plane Elements \u2202\u03be \u2202x = 1 a2 \u00d7 a = 1 a , \u2202\u03be \u2202y = 1 a2 \u00d7 (\u22120) = 0, \u2202\u03b7 \u2202x = 1 a2 \u00d7 (\u22120) = 0, \u2202\u03b7 \u2202y = 1 a2 \u00d7 (a) = 1 a . It should be noted that cases (a), (b) and (c) give the same results for the geometrical derivatives and the Jacobian. 5.2. Example: Influence of the shape regularity on the geometrical derivatives Given are two-dimensional elements as shown in Fig. 5.9. Calculate the geometrical derivatives of the natural coordinates (\u03be, \u03b7) with respect to the physical coordinates (x, y) and consider the different shapes as shown in Fig. 5.9a\u2013d. This problem relates to steps \u2776 to \u2779 as given on p. 256 and 257. 5.2. Solution The xy-coordinates of the four corner nodes for the different shapes are collected in Table5.7 5.3 Finite Element Solution 261 5.3. Example: Distorted two-dimensional element Given is a distorted two-dimensional element as shown in Fig. 5.10. Calculate the geometrical derivatives of the natural coordinates (\u03be, \u03b7) with respect to the physical coordinates (x, y). This problem relates to steps \u2776 to \u2779 as given on p. 256 and 257"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000645_jae-141897-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000645_jae-141897-Figure2-1.png",
        "caption": "Fig. 2. MG Generator for 2 MW class.",
        "texts": [
            " It will be shown that the MG Generator is able to obtain high speed operation, and its performance will be determined by using finite element analysis. \u2217Corresponding author: Tsubasa Oshiumi, Deparment of Adaptive Machine Systems, Graduate School of Engineering, Osaka University, 2-1 Yamadaoka, Suita, Osaka 565-0871, Japan. Tel./Fax: +81 6 6879 7553; E-mail: tsubasa.oshiumi@ ams.eng.osaka-u.ac.jp. 1383-5416/14/$27.50 c\u00a9 2014 \u2013 IOS Press and the authors. All rights reserved The proposed MG Generator has three major parts, as shown in Fig. 2: a low-speed rotor (LSR), a high-speed rotor (HSR) and a stator. The MG Generator\u2019s diameter and air gap length are almost the same as that of DD generators. The LSR is composed of 136 pole pieces made of laminated silicon steel sheets (50A470). The HSR has 32 neodymium iron magnets. The stator has 120 slots, 120 magnets and three phase distributed windings. First, the LSR which is directly connected to the rotor blades is rotated. Then, the HSR starts rotating synchronously according to the gear ratio which is determined by the following equations: Nl = Nh 2 +Ns (1) Gr = + Nl Nh/2 = 1 + 2 Ns Nh (2) Table 1 Combinations of MG generator No",
            " When the HSR rotates, it induces an EMF in the 3-phase windings, generating power. This combination of poles was chosen because it has a low cogging factor, Cf = 1. The cogging factor is defined by [5] Cf = NlNs LCM(Nl, Ns) (3) where LCM indicates the lowest common multiple. Possible pole combinations are given in Table 1. When the combination number is even, Cf becomes low and Gr is not an integer. To achieve low cogging torque, these topologies which will yield Cf = 1 are necessary. The MG Generator in Fig. 2(b) utilizes combination No.2 (Nh = 4, Ns = 15, Nl = 17, Gr = 8.5). The three phase windings are arranged as such in Fig. 3. When Gr is not an integer, the number of slots per pole per phase which is given by Eq. (4) becomes a fractional number. q = Ns 3 Nh (4) where q is the number of slots per pole per phase. When q is a fractional number, the windings become short-pitch windings. When the MG Generator is overloaded, it slips due to its inherent overload protection characteristics. The limit torque (slip torque) should be larger than the rated torque",
            "3 -2 -1 0 1 2 0 60 120 180 240 300 360 EM F [p .u .] Electrical angle [deg] U VW WU UV VW Fig. 6. EMF waveform. 0 500 1000 0 5.625 11.25 16.875 22.5 Tr an sm is si on to rq ue [k N m ] Rotation angle of HSR [deg] LSR HSR HSR: 120.1 kNm LSR: 120.1 kNm Gr8.5=1021 kNm Fig. 7. Transmission torque waveform. from sinusoidal waveforms. The THD of the phase and line voltage is 7.5% and 4.1% respectively. This is because of the rectangular shape of the laminated silicon steel sheets which are attached to the HSR magnets in Fig. 2(b). Since these laminated silicon steel sheets are rectangular blocks and not curved, the magnetic flux of the HSR at the air gap are square waves, which slightly increases the THD. In Fig. 5(b), torque as a magnetic gear and torque as a generator both contribute to the total torque. However, in order to evaluate only the torque as a generator, the two rotors should be rotated according to the gear ratio with the initial angle of the HSR being 0 deg, as shown in Fig. 5(a). The rated current of 6 Arms/mm2, which is almost the same value as conventional machines, was applied to the coils"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002541_ijsurfse.2016.076993-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002541_ijsurfse.2016.076993-Figure1-1.png",
        "caption": "Figure 1 A typical 2D view of the nozzle directed towards a substrate (see online version for colours)",
        "texts": [
            " This research work emphatically ponders on characterising the influence of scanning speed and high powder flow rate (PFR) on the volume of the deposited composites, evolving microstructure, porosity and the microhardness during laser metal deposition of Ti-6Al-4V/Cu composites. The scanning speed was varied for both the Ti-6Al-4V and Cu powders while the laser power and the powder feed rate were kept constant. The experiments were conducted at the National Laser Centre, Council of Scientific and Industrial Research (NLC-CSIR), Pretoria, South Africa. The equipment used was Ytterbium laser system equipment (YLS-2000-TR), powered at 2,000 Watts maximum and connected to a kuka robot with a three way powder jet nozzle attached at its end. Figure 1 shows a typical 2D view of the nozzle attached to the kuka robot. The nozzle is connected to a cylindrical component which embedded a glass of 2 mm thickness. The glass serves as a protector for the converging lens when the laser melts the powder. The rate of divergence at the focal point determines the spot size. The laser passes through the middle of the nozzle while the powders flow through the three way hole at the chamfered tip of the nozzle. The three way holes are at 120\u00b0 to each other and the powders were deposited concurrently onto the substrate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000863_s1061934814120168-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000863_s1061934814120168-Figure4-1.png",
        "caption": "Fig. 4. Effect of the pulse (a) amplitude and (b) time on the oxidation currents of cognac; supporting electrolyte, a phosphate buffer solution at pH 3.0; potential sweep rate, 10 mV/s.",
        "texts": [
            "59 V were observed on the differential pulse voltammograms of brandies and cognacs (Fig. 2). It can be seen that with pH increasing from 3 to 7, the potentials of peaks were shifted to the cathodic region, and their currents decreased; the first peak almost completely disap peared at pH 7.0 (Fig. 3). The best shape and charac teristics of the analytical signal were obtained at pH 3.0. The working conditions optimal for recording the analytical signal of beverages depending on the pulse amplitude and time were found (Fig. 4). The best results were obtained with a pulse amplitude of 50 mV and a pulse time of 50 ms. According to the Bureau National Interprofession nel du Cognac (Cognac, France), the main easily oxi dized compounds in cognacs are phenolic antioxi dants, extracted from oak wood when alcohols are matured in barrels [20]. These include ellagic and gal lic acids, syringaldehyde, vanillin, coniferaldehyde, furfural, and 5 hydroxymethylfurfural [10]. The potentials of their oxidation at an MWCNT\u2013GCE in the phosphate buffer at pH 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000057_ijsse.2019.100339-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000057_ijsse.2019.100339-Figure8-1.png",
        "caption": "Figure 8 Relation between the sensor protocol, the high criticality response times and the system\u2019s load status (see online version for colours)",
        "texts": [
            " It is only when the system becomes overloaded that it becomes unstable and response times start rising fast. Moreover, a spike exits in the requests\u2019 response time after the overload of the system. That is due to leftover low-level requests, that have arrived before the overload of the system, and could not be handled until after the system was back to normal status. The second presented policy, regards the way the sensors communicate. The policy chooses between a push or pull communication model. The results of this policy are presented in Figure 8. In the top chart, the pushing or pulling mode of the sensor is depicted by the background of the chart; the dark background indicates pushing while the lighter one pulling. Additionally, the average response time for high criticality requests as well as the per-interval response time for high criticality requests is shown. On the bottom chart, the number of active cores along with the number of overloaded cores is presented. When the sensor turns to pushing mode, the average response time drops significantly, and only increases as the system overloads and becomes unable to handle the load"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000497_9781119011804.ch6-Figure6.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000497_9781119011804.ch6-Figure6.10-1.png",
        "caption": "FIGURE 6.10 Optical cell stretching. (a, b) Cell stretching using attached microbeads. Adapted from Reference 67. Two 4.12-\u03bcm diameter silica microbeads are nonspecifically attached to the red cell at diametrically opposite points. (a) The left bead is anchored to the surface of the glass slide. (b) The right bead is trapped using the optical tweezers while the slide and attached left bead are moved to stretch the cell. The image shows deformation at 193 pN of force. (c) Optical stretcher: counter-propagating beams trap and stretch the cell. Adapted from Reference 68. Mechanics and chemistry biosystems. Not drawn to scale; optical fiber diameter is 125 \u00b1 5 \u03bcm.",
        "texts": [
            " At the cell level, the view that deformation and other mechanical properties of cells can give insights on possible connections with cellular process such as cell differentiation, cell health, and disease progression is driving multidisciplinary research initiatives at the interface between physical science and biology3. Different mechanical models of the cell have been developed [66] based on the different experimental techniques that were used for the mechanical measurements. Two examples of using optical traps for investigating the viscoelastic properties of cells are illustrated in Figure 6.10. The procedure by Mills et al. [67] is illustrated in Figures 6.10a and 6.10b, where a red blood cell is stretched by attaching beads to the cell, pulling on the bead, and then analyzing the cell deformations. The out-of-focus bead on the right is held in place by a calibrated optical trap as the in-focus bead on the left, fixed to the bottom of the chamber is displaced using a translation stage. Figure 6.10c illustrates the so-called optical stretcher by Guck et al., which uses counter-propagating beams to directly trap and stretch the cell. The optical stretcher mechanism can be intuitively understood in terms of the refractive index dependence of the optical momentum flux, nmP\u2215c, where nm is the refractive index of the medium 3For example, the U.S. National Cancer Institute runs a program \u201cPhysical Sciences in Oncology\u201d aimed at building cross-disciplinary teams working at the convergence between cancer biology and physical science/engineering"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001030_haptics.2014.6775481-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001030_haptics.2014.6775481-Figure1-1.png",
        "caption": "Figure 1: Concept of the haptic enhanced reality",
        "texts": [],
        "surrounding_texts": [
            "Haptic rendering technologies have long been proposed as a function that interfaces should have for the intuitive transmission of information from a system to a user. There are many studies in this field such as tele-operation[?] and rehabilitation[?] reflecting a strong need for haptic rendering. Methods for calculating force responses in the deformation of an object have also been studied [?, ?]. Some studies have reported the haptic information is very useful in training surgical skills [?, ?]. Mechanical properties of a variety of human organs and tissue have been measured [?], and these properties were used to calculate the force response when manipulating virtual human tissue [?]. In these studies, highprecision modeling of the force response requires significant calculation time and high-performance devices to fully render the haptic senses. With poor performance of the model or the device, users feel a sense different from the one they should feel. Our research group have proposed the use of force rendering by augmentation to realize a realistic force response even using a lowperformance haptic device [?]. The key idea is to use an object that has material properties similar to those of a target object, and to overlap the force produced by the haptic device on the force of the base object as shown in Fig. ??. The haptic senses produced by the proposed Haptic Enhanced Reality (HER) method are improved because the base object covers a haptic response that we cannot perfectly model or that the device cannot completely produce. In this paper, we explore the feasiblity of the HER method to a needle insertion task. The needle insertion task is a basic manipulation for the manual skill training in the fields of nursing and medicine. We first explain the outline of the HER method and provide a general discussion of the characteristics that the base object should have. Next, we explain the method to change the elastic property and the tear resistance of a rubber sheet by stretching. The force response of the rubber sheet can be controlled by changing the \u2217e-mail: kurita@bsys.hiroshima-u.ac.jp intensity of stretching. We also describes the experimental results to confirm the efficacy of the proposed method."
        ]
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure6-1.png",
        "caption": "Figure 6 Schematic diagram that shows the description of cutter run out.",
        "texts": [
            " Furthermore, considering the fact that pitch deviation can be both positive and negative values, and that it is a kind of manufacturing error, one can assume that it follows a normal distribution of mean zero. Therefore, in the model which is explained later, pitch deviation was represented as the cutter rotation deviation that follows normal distribution of mean zero. In an actual skiving, the cutter shaft is long as 30mm while the run out of the cutter is small as 20 m. Therefore, the posture variation of the cutting edges can be ignored. For these reasons, the cutter run out can be described as a revolution around its original axis with its orbit radius equal to the half the run out (Fig. 6). 3 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/86626/ on 03/02/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Three coordinate systems Ow-xwywzw, Oc-xcyczc and Of-xfyfzf were established as shown in Fig. 7. Ow-xwywzw and Oc-xcyczc were attached to the workpiece and cutter respectively, and rotate with them. Of-xfyfzf is connected to fixed frame. The cutting edge of the cutter was determined by calculating the conjugate tooth flank of the internal gear, and was fixed to the cutter coordinate (i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003062_s1068366616050044-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003062_s1068366616050044-Figure2-1.png",
        "caption": "Fig. 2. General diagram of tribometer.",
        "texts": [
            " In this study, EXFS of ElvaX Company were used for the express analysis of elemental composition and concentration of friction modifiers in neutral base oil. EXFS from ElvaX allowed the detection of elements from sodium (atomic number is 11) to uranium (atomic number is 92) by recording the spectra at the emitter voltage of 40\u201349 kV. For the exact evaluation of a composition that does not contain light elements, 10 s were sufficient. The resolution is up to 200 eV by the line of 5.9 keV (Fe isotope). The 4096-Channel analog-to-digit transducer was used. CSM tribometer (Fig. 2) was used to determine the volume wear and coefficient of friction in online mode at various temperatures, contact pressures, speed, and moisture content in the lubricating medium. The measurement principle is given in Fig. 3; spherical indenter (sphere) was dropped to the studied counterbody at known precise load. Counterbody represented by disc was mounted on an elastic lever connected to the friction force sensor. During the rotation of the disc sample between counterbody and indenter, the friction force arose, which was measured by minimum deviation of elastic lever using LVDT sensor (inductive sensor of linear displacements)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003195_978-981-10-2875-5_2-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003195_978-981-10-2875-5_2-Figure2-1.png",
        "caption": "Fig. 2 The D-H coordinate system of a simulated manipulator platform",
        "texts": [
            " The 3-D view of simulated manipulator platform is shown in Fig. 1. Due to the kinematic model of hyper-redundant manipulator should not be affected by the shape or arrangement of the links change, in this work, we simply it into a linkage mechanism. The kinematic model of the manipulator has 14 DOFs joints with each joint has p=2 range of motion. The method that is implemented here for modelling forward kinematic of the manipulator is D-H method [9]. The D-H coordinate system of a simulated manipulator platform is shown in Fig. 2. The D-H parameters are listed in Table 1. Where Ai is the link corner of hyper-redundant manipulator, ai is the length of vertical line from the joint shaft i to i + 1, di is the link offset from link i to i + 1, hi is the angle of rotation of link i. According to these D-H parameters of the manipulator, it is possible to establish the transformation matrix between coordinate frames i and i + 1. The matrix can be given as: i 1 i T \u00bc Rot\u00f0x; ai 1\u00deTrans\u00f0ai 1; 0; 0\u00deRot\u00f0z; hi\u00deTrans\u00f00; 0; di\u00de \u00f01\u00de where Rot () is the rotation coordinate transformations, Trans () is the translation coordinate transformations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002253_ecce.2014.6953966-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002253_ecce.2014.6953966-Figure5-1.png",
        "caption": "Fig. 5. Geometry of the first inset machine",
        "texts": [
            " Moreover, the rotor position affects the saturation as well so that also the parameters rely on it. As a first example, an inset permanent magnet (PM) machine is here considered. The inset machine is derived from a surface mounted PM machine with additional iron teeth in the rotor between each couple of magnets. Such teeth introduce a magnetic anisotropy detectable in any operating conditions. Nevertheless, due to the nonlinear characteristic of the iron, the self-sensing capability is expected to vary with the rotor position and with the steady-state current [14]\u2013[16]. Fig. 5 shows a sketch of the machine geometry. It is a 12-slot 8-pole machine characterized by a fractional-slot nonoverlapped winding [17], [18]. During the analysis eddy current in the rotor are included in the finite element simulations. The effects of the rotor position on the saturation of the lamination is shown in Fig. 6, where the flux density distribution is reported for two different rotor positions. As can be noted, the saturation of the iron is quite different. In particular the magnetic load is very different in the stator teeth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000238_ijvd.2019.101524-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000238_ijvd.2019.101524-Figure1-1.png",
        "caption": "Figure 1 The working principle of transfer case",
        "texts": [
            " When most of the lubricating oil in the gap between the friction plate and the dual steel disc is extruded, the bearing capacity and torque of the oil film no longer exist, which means the mixed friction stage is over. Stage 3: Asperity torque. The pressure is completely carried by the asperity contact, and the mechanical torque generated by the asperity torque constitutes the friction torque of the engagement process. This state will last until the relative speed of the friction plate and the dual steel disc is zero. The working principle is shown in Figure 1. In these three stages, the multi-plate clutch will inevitably consume power and generate a large amount of heat in the process of frequent and high-speed engagement, which is the working characteristics of the multi-plate clutch. In the process of sliding, the temperature distribution of each part of the friction pair is uneven, resulting in thermal expansion in friction and dual steel disc. And the excessive temperature of the friction pairs will cause thermal stress and deformation. Once the friction plate or dual steel disc internal stress value is beyond the yield point of the material, irreversible plastic strain will occur, and the friction heat that mainly stays on a certain deformed surface will make the temperature soar, producing thermal crack formation, friction vibration and other adverse conditions. Therefore, it is necessary to study the torque transfer characteristics of friction plate and dual steel disc under the thermal load disturbance. The working principle of the multi-plate clutch is shown in Figure 1. In the process of sliding friction of the multi-plate clutch, when the friction plate and the dual steel disc are assumed to be two solids, the heat generated per unit area per unit time is related to friction coefficient, contact pressure, sliding angular velocity and radius. To calculate the friction torque of clutch, a micro ring with a thickness of dr is taken from the friction surface of the clutch, and the radius of the ring is r . So the area of micro ring is 2dA rdr\u03c0= , and the friction torque generated on the ring is 22app appdT r p dA r p dr\u03bc \u03c0 \u03bc= = "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002884_1.4034511-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002884_1.4034511-Figure3-1.png",
        "caption": "Fig. 3 LMMB raceway",
        "texts": [
            "5lm, with flatness value of 0.2 mm. The flatness was verified with high precision spirit meter. Heat flow information during testing was measured by 24 (k-type) thermocouples. There are three thermocouples inserted in the face of each specimen by putting 3 mm diameter thin holes. Remaining four is fixed near corner and one is fixed at the center of the specimen. Similarly, two sets of LMMB were manufactured to carry out the temperature distribution studies. The detailed LMMB manufacturing diagram is shown in Fig. 3. The axis of the groove would be an arc of a circle about the axis of rotation. The rolling elements (balls of standard sizes 31.75 mm) were purchased from commercial manufacturers made of through hardened (62 HRc) high-carbon chromium bearing steel (AISI 52100). The sections of complete raceways were represented by small rectangular steel blocks [17\u201319]. Each block had a cylindrical groove on one face, parallel to opposite face, having a radius slightly greater than that of the ball with which it was to be used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002736_j.mechmachtheory.2016.07.025-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002736_j.mechmachtheory.2016.07.025-Figure2-1.png",
        "caption": "Fig. 2. Six-bar mechanism and equivalent four-bar linkage-cam mechanism in dwell configuration.",
        "texts": [
            " The basis for such a substitution is the theorem on coordinated centers in envelope theory. The theorem as derived by the previous researchers cannot be used for substitution when the relative motion is at a T1-position of the first-kind. Because of the present extension of the classical result to T1-position of the first-kind, the usual substitution procedure readily applies here too. We show an example of equivalent mechanisms where the substitution is applied between two links whose relative motion is in a T1-position of the first-kind. Fig. 2 shows a six-bar Stephenson-III mechanism in a dwell configuration. The center of curvature of the instantaneous point-path of coupler point G coincides with the joint center Go. Consequently, link 6 will be in a second-order instantaneous dwell configuration. In this particular configuration of the mechanism, the four-bar sub-chain EoEFFo is an asymptotic configuration [18] which means that the coupler link 3 is at a T1-position of first-kind with respect to the ground link. Since link 6 is in a dwell up to second-order instantaneous kinematics, this link can be considered as ground link instantaneously"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000534_9781118889664.ch25-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000534_9781118889664.ch25-Figure1-1.png",
        "caption": "Fig. 1. Finite element model used in the simulations.",
        "texts": [
            " Then the temperature history was used as input to mechanical model to calculate the thermal stress evolution during the process and the residual stress after the sample completely cooled down to room temperature. The influences of process parameters including the scanning speed and scanning strategies on temperature history, thermal stress evolution, and residual stress were investigated in details. In this investigation, a three dimensional finite element model was constructed to simulate the thermo-mechanical behavior of laser cladding process based on ABAQUS. Fig. 1 shows the geometry and mesh of the finite element model. The coating layer was built by scanning 10 adjacent single tracks, each with a length of 10.0 mm, a thickness of 0.25 mm and a width of 1.0 mm. The coating layer was fabricated on the surface of a substrate having 5 mm thick, 10 mm wide and 20 mm long. A dense mesh was employed for the coating layer and the contact area with the substrate since higher thermal gradients exist in these regions. Three different scanning speeds (2, 5, and 10 mm/s) and two different scanning strategies (alternative Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002279_detc2015-46402-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002279_detc2015-46402-Figure4-1.png",
        "caption": "Figure 4 Schematic diagram of the effect of the axial feed rate on skived gear accuracy.",
        "texts": [
            " Therefore, in order to exclude such factor, material that has low cutting resistance such as carbon was employed as the material of the blank. Table 2 shows the pinion-type cutter data and Fig.3 shows a photograph of the cutter. The cutter accuracy was of class JIS AA. To generate internal gear teeth through whole face of the workpiece, a cutter with synchronous rotation must be fed in the direction of the axis of the workpiece. Therefore, spiral cutting marks are formed on the tooth flanks, and its pitch increases in proportion to the feed rate, which deteriorates the gear accuracy as shown in Fig. 4. However, because internal gear and cutter have a high contact ratio, cutter with pitch deviations, directly influences pitch deviations of the skived gear. Now let\u2019s focus in one specific tooth space of a gear. Pitch deviation is the difference of distance between adjacent tooth flank. Because in our model, simulations are conducted in only one tooth space of the workpiece, pitch deviation can be described as a rotational deviation of the cutter (Fig. 5). Furthermore, considering the fact that pitch deviation can be both positive and negative values, and that it is a kind of manufacturing error, one can assume that it follows a normal distribution of mean zero"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001972_ilt-03-2015-0034-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001972_ilt-03-2015-0034-Figure1-1.png",
        "caption": "Figure 1 Schematization of plain journal bearing",
        "texts": [
            " These analyses are carried out by governing continuity equations consisting the Navier-Stokes, turbulent kinetic energy and its dissipation rate equations coupled with the displacement equation for the bush are solved to obtain the three-dimensional steady-state EHD characteristics of plain cylindrical journal bearings. The lubricant flow in turbulent regime is modeled using the AKN low-Re k 2 e turbulence model. The pressure field is determined by the resolution of the generalized Navier-Stokes equation under classic assumptions in the (O, \u00a1, z\u00a1) coordinate system. Figure 1 illustrates the schematization of plain cylindrical journal bearing. This equation can be expressed by the following form [10]: ( \u00a1 U) 0 (1) Where \u00a1 U \u00a1 U (u, v, w) is the velocity vector. Equation (1) can also be written as follows: u x v y w z 0 (2) The Navier-Stokes equation can be defined in the following form (1997): \u00b7 (U \u00a1 U \u00a1 ) p . ( U \u00a1 ( U \u00a1 )T) B (3) With, P static pressure (thermodynamic); U velocity; dynamic viscosity. For fluids in a rotating frame with constant angular velocity , source term B can be written as follows: B (2 \u00a1 U \u00a1 \u00a1 ( \u00a1 r\u00a1)) (4) Equation (1) can also be expressed in the form: u u x v u y w u z p x 2 x2 u 2 y2 u 2 z2 u BX (5) u v x v v y w v z p y 2 x2 v 2 y2 v 2 z2 v BY (6) u w x v w y w w z p z 2 x2 w 2 y2 w 2 z2 w BZ (7) is fluid density"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003518_j.engfracmech.2015.05.004-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003518_j.engfracmech.2015.05.004-Figure13-1.png",
        "caption": "Fig. 13. Direction of tangential stresses in (a) an accelerating wheel; (b) a braking wheel.",
        "texts": [
            " 11 the crack with a 30 inclination is found to be the most severe case. For this crack, significantly increased displacement magnitudes at each loading cycle (i.e. ratcheting) are observed. Tearing displacements are found to be comparably small also in position P2 (see Fig. 11(d)). (see Fig. 12). In [3], it was observed that deformations, stress and plasticity levels in-between short surface cracks subjected to tractive rolling with f \u00bc 0:3 are larger than for f \u00bc 0:3. Directions of tangential stresses in an accelerating and a braking wheel are shown in Fig. 13. As for peak crack tip shear displacements this difference is not very pronounced for the radial and the 50 inclined crack as seen in Fig. 14(a) and (b). However, for the 30 -crack the crack tip shear displacement is significantly larger and the ratcheting behaviour more pronounced for f \u00bc 0:3 than for f \u00bc 0:3, as observed in Fig. 14(c). The elastoplastic response corresponding to 3D rolling contact featuring full slip is now compared to the response under partial slip conditions. The computational scheme used for implementing rolling contact loading under partial slip conditions is detailed in [1]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003349_indin.2016.7819202-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003349_indin.2016.7819202-Figure2-1.png",
        "caption": "Fig. 2. Park\u2019s vector transform",
        "texts": [
            " Principal component analysis Id Iq v Fuzzy logic Park\u2019s vector transform Obtainin feature signal Rule base Inference mechanism TB2 Stator condition TB1 Ia Ib Current signals Ic Fig.1. Fuzzy logic based fault detection method CBq CBAd iiI iiiI 2 1 2 1 6 1 6 1 3 2 \u2212= \u2212\u2212= (1) In (6), CBA iandii ,, represent phase currents, respectively. Under ideal conditions, three phase current signals constitute the park\u2019s vector component by using (2). ) 2 sin( 2 6 sin 2 6 \u03c0\u03c9 \u03c9 \u2212= = tiI tiI Mq Md (2) In (2), in and \u03c9 represent the maximum value of supply current and supply frequency, respectively. The transformation made on three phase current signals are given in Fig. 2. In Fig.2, park\u2019s vector transform construct a circular pattern at the center of two components. The faulty conditions are detected by monitoring the deviations in the pattern of this transform. Whereas a two dimensional pattern formed by park\u2019s vector transform is a circular shape in a healthy motor, it turns into an elliptical shape in a faulty condition. B. Principal Component Analysis Principal component analysis is a statistical data analysis. This method is used for new components from an original data set [24]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001784_gt2015-43638-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001784_gt2015-43638-Figure4-1.png",
        "caption": "Figure 4. Velocity magnitude contours for representative line-to-line (a,b,c) and clearance (0.005in.) (d,e,f) cases.",
        "texts": [
            " This pressure distribution pattern is also weakly seen for a higher backing plate fence when Fig. 3a and 3c are compared. The wider opening at the backing plate fence region eliminates pressure drop accumulation by allowing more leakage with relatively lower contraction effect. These relatively uniform pressure contours for a thicker bristle pack or higher backing plate fence do not indicate that there is no inward radial flow. They indicate that the inward radial flow is less pronounced as quantified in the velocity plots in Fig. 4(a,b,c), 5(a,b,c), and 6(a,b,c). The velocity vector plots in Fig. 5(a,b,c) and 6(a,b,c) show that the flow diffuses into the bristle pack from the upstream face. The diffusing flow into the bristle pack in the radial bristle free height accumulates on the backing plate. The flow facing the backing plate radially directs through the backing plate fence since this region is the only opening region to the downstream side. This flow pattern is in the inward radial direction and it forms the driving force that cause bristle blow-down towards the rotor surface",
            ": The flow configuration for the clearance cases differs from that of the line-to-line cases due to the clearance dominated flow pattern. First of all, the flow patterns are compared for all the clearance cases for the baseline case and cases at maximum bristle pack thickness of 0.100in. and maximum backing plate fence height of 0.100in. as plotted in Fig. 3-6(d,e,f). The clearance is taken as 0.005in. for Fig. 3-6(d,e,f). Velocity magnitude contours are plotted by zooming at the fence height region in order to better visualize the clearance region in Fig. 4(d,e,f). Fig. 3(d,e,f) compares the effects of bristle pack thickness on the pressure contours. As is similar to the line-to-line case, the pressure drop is more evenly distributed over the bristle pack for a thicker bristle pack and a higher backing plate fence. This effect is clearly seen in Fig. 8, which shows the radial pressure distribution over the backing plate for both clearance of 0.005in. and 0.010in. The inward pressure drop is very high near the backing plate fence region for the baseline case when compared to the thicker bristle pack or higher backing plate fence",
            " 8, the radial pressure over the backing plate is also strongly dependent on the clearance level. With increasing clearance, radial pressure drop level increases, meaning that the pressure around the bristle pinch region is getting closer to upstream pressure. At higher clearances, the pressure drop occurs around the bristle fence region. The deviations between the present CFD and others [5,12] are due to different brush seal geometries, especially different bristle pack thickness as pointed out above for the lineto-line cases. As seen in the velocity plots of Fig. 4(d,e,f), 5(d,e,f),and 6(d,e,f) the main leakage flows through the clearance while the flow diffusion into the bristle pack still occurs as in the line-toline case. The location of maximum velocity for the clearance case is at the clearance downstream side, while it is at the inner edge of the backing plate for the line-to-line case. Among clearance cases, the flow velocity patterns are not considerably affected by the bristle thickness and the backing plate fence height due to clearance dominating flow formation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001214_ijhvs.2014.068101-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001214_ijhvs.2014.068101-Figure1-1.png",
        "caption": "Figure 1 Plan view of the wheelset on inertial X-Y plane",
        "texts": [],
        "surrounding_texts": [
            "Two views of a wheelset on a curved track are shown Figures 1 and 2. Coordinate frame XYZ is the fixed frame and xyz is gimble (moving) frame with its origin at the centre of mass G. The wheelset spins with angular speed \u03c9 about the y-axis. Rotation of the moving frame is defined by \u03b8 and \u03c8 about x-axis and z-axis, respectively. \u03c6 is the component of \u03c8 along inertial Z-axis and \u03c6 is the angle made by the x-axis with respect to inertial X-axis in the horizontal plane. Total angular velocity of the wheelset is defined by its components ,\u03b8 \u03c9 and \u03c8 about the axes of the moving frame. \u2018a0\u2019 and \u2018b0\u2019 are the centres of instantaneous rolling circles. The two wheel-rail contact points are \u2018a\u2019 and \u2018b\u2019 and are henceforth termed as contact \u2018a\u2019 and contact \u2018b\u2019, respectively. If Ixx, Iyy and Izz are the principal moments of inertia of the wheelset about x, y and z axes, respectively, and Mx, My and Mz are the moments about the axes of the gimble (moving) frame then rotational dynamics of the wheelset is defined by the following equations: xx x yyI M I\u03b8 \u03c9\u03c6= + (1) yy yI M\u03c9 = (2) .zz z yyI M I\u03c6 \u03c9\u03b8= \u2212 (3) The equations for translational dynamics are: w g X w g Y w g Z m X F m Y F m Z F = = = (4) where FX, YF and FZ are the resultants of contact and external forces along the axes of the inertial frame XYZ and mw is the mass of the wheelset. Xg, Yg and Zg are the coordinates of the centre of the mass of the wheelset."
        ]
    },
    {
        "image_filename": "designv11_64_0001901_s1068798x15080055-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001901_s1068798x15080055-Figure1-1.png",
        "caption": "Fig. 1. Spiroid gears: (a) quarter turn gear; (b) multiturn gear; 1, housing; 2, base; 3, input flange (lid); 4, adapter; 5, safety valve; 6, spiroid screw; 7, gear.",
        "texts": [
            " They are used at working temperatures between \u201360 and +50\u00b0C, under heavy loads, and at high speeds. Consequently, their efficiency is relatively low and the wear rate of the frictional pairs is high. Such gear systems must have long life. Therefore, it is important to find means of improving their performance\u2014for example, reducing the frictional forces in the slip bearings by the appropriate selection of the gear materials and lubricants. We investigate the RZA S 2000 quarter turn gear for a pipe rolling system (Fig. 1a) and a multiturn RZAM S 1000 gear (Fig. 1b). The quarter turn gears are mounted on ball valves and disk gates. They are characterized by brief operation per cycle (no more than 200 s); the working component of the rolling system rigidly connected to the spiroid gear rotates by 90\u00b0 \u00b1 10\u00b0. The multiturn gears are used to con trol gate valves and similar equipment. Their working period is longer (5\u201360 min per cycle, 40\u2013150 turns of the gear for 1000 or more turns of the gear\u2019s input shaft). The characteristics of the spiroid gears considered are as follows: RZA S 2000 RZAM S 1000 Interaxial distance aw, mm 60 40 Gear ratio i12 46 11.46 Axial module of spiroid screw mX, mm 2.750 2.774 External diameter of spiroid screw da1, mm 42 48 Internal/external diameter of spiroid gear di2/de2, mm 138/175 120/155 Maximum torque on gear\u2019s output shaft T2max, N m 2000 1000 Limiting static torque on output shaft T2li, N m 4000 2000 Mass m, kg 17 11 The gears (Fig. 1) include a housing 1 that contains the spiroid screw 6; input flange (lid) 3; adapter 4 (the output shaft for the quarter turn gear) or a protective dome (for the multiturn gear); and a transmission consisting of the spiroid screw 6 rotating in roller bear ings, and the gear 7, rotating in slip bands. Auxiliary com ponents include indicators, speed limiters, and structural components. The RZA S 2000 system (Fig. 2a) is based on spiroid gear 1, radial bearings 2, and an end bearing 3, while the RZAM S 1000 system (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002524_j.procir.2016.02.049-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002524_j.procir.2016.02.049-Figure3-1.png",
        "caption": "Fig. 3. Tool path during process step (surface milling)",
        "texts": [
            " This is as high as possible to improve ductility, without exceeding the point above which embrittlement occurs ( \u2248 980\u00b0C). The tool temperature was set at 20\u00b0C and the punch velocity at 20 mm/s. 3.2. High performance machining Process I Process I includes face milling and swarf milling. The cutting parameters strongly depend on the tool used and the geometric characteristics of the machined features as summarized in Table 1. The strategy therefore not only refers to the toolpath, but the specific combination of the toolpath with corresponding cutting parameters. A toolpath on the flange is shown in Fig.3 Process II The high performance cutting strategy implemented for Process II included a constant engagement angle (CEA) milling strategy. With the geometric features considered, the optimal engagement angle is calculated. From this, the tool paths are created so that the engagement angle is controlled according to the features surrounding the immediate toolpath. This is of particular importance for titanium to reduce cutting forces and increase tool life [4]. Finishing strategies included swarf milling and surface milling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001594_aim.2014.6878295-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001594_aim.2014.6878295-Figure14-1.png",
        "caption": "Figure 14. Experimental setup.",
        "texts": [
            " We developed the walking support robot equipped with four velocity-based mechanical safety devices and two torque-based mechanical safety devices. Fig. 13 shows the developed robot. As shown in Fig. 13, the length is 125[cm] and the width is 154[cm]. The armrest is adjustable in height from 85[cm] to 108[cm] according to the height of each patient by using a hand crank. We conducted the following experiments by using the developed walking support robot. We experimentally examined whether the torque-based safety device can achieve the function. Fig. 14 (a) shows the experimental setup. We mounted the walking support robot on the two spacers. Then, we attached a steel bar to the right wheel and brought the steel bar into contact with the block. After that, we measured the current supplied to the right drive unit\u2019s motor while increasing the motor torque. The sampling frequency was 200[Hz]. We experimented using three detection torque levels of 10.0, 15.0, and 20.0 [Nm]. Ten trials were conducted for each detection torque level. The detection velocity level was 1",
            " The differences between the detection torque levels and the experimental values are attributed to the repeatability errors of the torque limiter, the attachment position errors of Adjusting Nut, the output torque errors of the motor, among others. B. Velocity-based Safety Device Next, we removed the steel bar from the right wheel and attached some markers on Gear A (i.e. Shaft C), Claw D, and Plate A of the velocity-based safety device 1. Then, we measured the velocity of Shaft C, the motion of Claw D, and the motion of Plate A by using a motion capture system (HAS-500, DITECT Corporation) while increasing the velocity of Shaft C by the motor (see Fig. 14(b)). The sampling frequency of the motion capture system was 200[Hz]. We experimented using two detection velocity levels of 1.00 and 1.50 [rad/s]. The detection torque level was 10.0 [Nm] in the torque-based safety device 1. Ten trials were conducted for each detection velocity level. Fig. 16 shows the typical example of the experimental results. In Fig. 16, the time when Claw D locked Plate A is indicated by an arrow. Fig. 16 indicates that the velocity of Shaft C was approximately the detection velocity level at the time when Claw D locked Plate A"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003524_tdcllm.2016.8013240-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003524_tdcllm.2016.8013240-Figure7-1.png",
        "caption": "Fig. 7. Identification of directions in case of different shapes of connecting elements",
        "texts": [],
        "surrounding_texts": [
            "One of the most important questions regarding to the mechanical loading conditions of any structure is the distribution and the maximal value of forces affecting to the different nodes. Fig. 8 shows the distribution of the forces in the model. The maximal force value was 611.48 N(y) in case #1, 754.86 N(y) in case #2 and 1718.55 N(y) in case #3. It can be determined that asymmetry has a notable effect on the magnitude of maximal forces affecting to the structure. In case of symmetric loading conditions, none of the partial forces reached value of the total mechanical load (1500 N)."
        ]
    },
    {
        "image_filename": "designv11_64_0000702_1.4031894-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000702_1.4031894-Figure6-1.png",
        "caption": "Fig. 6 Schematic deformation of annulus during (a) prebuckling (axisymmetric) and (b) postbuckling (periodic). (a) Conservation of radial lengths in the initial planform suggests that the outer edge of the deformed plate must contract, leading to circumferential compression; the variation overall must be self-equilibrating, so there is in tension on the inside and a nonuniform variation in between, as confirmed by finite element analysis. (b) Subsequent buckling into a periodic mode shape of wavelength, k: thus, n 3 k 5 2pa.",
        "texts": [
            " The exact variation of strains is found by solving the pair of F\u20acoppl\u2013von Karman governing equations of deformation [7] but tractable solutions are rare, even for simple geometries, and semi-analytical schemes are usually pursued. The previous finite element analysis offers complementary insight and confirms that the in-plane stress field is increasingly dominated by circumferential stresses, rh, as b/a increases, with compression on the outer edge and tension on the inner edge. If we imagine that the smaller radial strains are, in fact, zero, the outer edge deflects transversely and inwardly to preserve the radial length. Consequently, the outer edge length must contract and hence, compress, see Fig. 6. The inner edge must be in tension at all times so that the resulting radial variation is self-equilibrating, i.e., there is no nett circumferential force for none is applied externally. These properties are confirmed in Fig. 7, which shows some finite element data for the prebuckling phase for different initial widths: rh is always compressive on the free edge, and its variation across the annulus tends more toward being linear as the annulus width decreases. The annulus therefore buckles because of unsustainable compression on the free edge, giving way to a periodic \u201crippling\u201d that superposes onto the prebuckling deflections cf"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.23-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.23-1.png",
        "caption": "Fig. 4.23 External and internal forces in a mechanical system",
        "texts": [
            " However, not all types of forces are relevant in the multibody systems of common application. Figure 4.22 shows a mechanical system composed by six links and a free sphere that can collide with the slider (body 6). In this system, it is possible to identify the main types of forces described above. In a broad sense, the forces in the multibody systems can be divided into two main groups, namely external and internal forces. If the forces are located inside of the boundaries of the system, they are called internal, otherwise the forces are named external forces. Figure 4.23 illustrates a MBS in which the internal and external forces can easily be identified. The definition of the system boundaries depend on the system in analysis. Figure 4.24 illustrates a body i acted upon by a gravitational field in the negative y direction. The choice of the negative y direction as the direction of gravity is totally arbitrary. However, in the present work, the gravitational field will be considered to be acting in this direction unless indicated otherwise. The force due to gravitational field can be written as FG \u00bc mig~uy \u00f04:93\u00de where mi is the mass of body i and g is gravity acceleration"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000497_9781119011804.ch6-Figure6.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000497_9781119011804.ch6-Figure6.8-1.png",
        "caption": "FIGURE 6.8 Setup for single-cell laser tweezers Raman spectroscopy (LTRS). Adapted from Reference 42. PNAS. (a) Brightfield images of individual trapped microalgal cells: 1, Neochloris oleoabundans; 2, Botryococcus braunii; and 3, Chlamydomonas reinhardtii. (b) Instrument layout. The 785-nm laser beam is used as both trap beam and Raman excitation beam. (For a color version of this figure, see the color plate section.)",
        "texts": [
            " Aside from temperature control, microfluidics also provides the possibility for introducing new cells and varying the chemical microenvironment of the trapped cell [41]. One application of environmental control is in real-time control of cellular growth conditions. One area of interest is in finding optimal growth condition for producing lipid in algae since such lipids can be readily converted to biodiesel. In an effort to establish correlations among lipid trigger conditions, growth rate, and lipid production in algae, Wu et al. used LTRS, illustrated in Figure 6.8, to quantitatively investigate the lipid content of trapped single cells from different algal species and different growth conditions including light exposure, nitrogen, silicon, CO2, pH, and temperature [42]. Differences in temporal response of individual cells, such as lags in the onset of cellular response, can cause averaged bulk measurements that paint a different picture of the cell population. Using optical traps to hold cells can keep them in the observation zone for monitoring temporal effects after their chemical microenvironment is modified, for example, by pumping fluids containing different reagents"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-Figure3.44-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-Figure3.44-1.png",
        "caption": "Fig. 3.44 Simply supported Bernoulli beam under pure bending load",
        "texts": [
            "13 Simply supported beam with centered single force: analytical solution Calculate the analytical solution for the deflection uz(x) and uz ( L 2 ) of the simply supported Bernoulli beam shown in Fig. 3.43 based on the fourth order differential equation given in Table3.5. It can be assumed for this exercise that the bending stiffness EIy is constant. 174 3 Euler\u2013Bernoulli Beams and Frames 3.14 Simply supported beam under pure bending load: analytical solution Calculate the analytical solution for the deflection uz(x) and uz ( L 2 ) of the simply supportedBernoulli beam shown in Fig. 3.44 based on the second order differential equation for the bending moment distribution given in Eq. (3.39). It can be assumed for this exercise that the bending stiffness EIy is constant. 3.15 Bernoulli beam fixed at both ends: analytical solution Calculate the analytical solution for the deflection uz(x) and slope \u03d5y(x) of the Bernoulli beams shown in Fig. 3.45 based on the fourth order differential equation given in Table3.5. Determine in addition the maximum deflection and slope. It can be assumed for this exercise that the bending stiffness EIy is constant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001271_amm.611.279-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001271_amm.611.279-Figure2-1.png",
        "caption": "Fig. 2 The model of gear with variable transmission",
        "texts": [
            " To determine the stress in the foot the tooth for gear with asymmetrical profile this approach is insufficient. One way to determine the stress in a dangerous section of the tooth is by finite element method. This problem is solved for gear with variable transmission in the range u = 0,5 to 2,0 ,with the number of teeth z1 = z2 = 24 and gearing module mn = 3,75 mm, the axial distance a = 90 mm and for a one sense of rotation. To create this gear is analyzed in detail in the literature [5] and [6]. The gears for a given variable transmission have been proposed as elliptical - eccentrically placed (Fig. 2), so that conditions were right shot. Real of load gear teeth with variable gear ratio is not constant. By way of illustration is given unit input torque (driven) spur gear Mk1 = 100Nm. In Figure 3 the course of torque Mk1 on the input gear and torque Mk2 on the output (driven) gear (Mk2i = Mk1.ui) is show. In Figure 4 are value of changing tangential tooth load the driver and driven gear F01 = F02 if F01=Mk1/r1i, radial force Fr1 = Fr2 if Fr1 = F01.tg\u03b1 . The resultant force acting on the side of the tooth FN1 = FN2 if FN1 = F01 / cos\u03b1, where \u03b1 is an angle of action to 20\u00b0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001350_15325008.2014.943438-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001350_15325008.2014.943438-Figure2-1.png",
        "caption": "FIGURE 2. The action process of EMA.",
        "texts": [
            " The trip coil excited a current to magnetize the movable axle center, which formed an electromagnet that generated a magnetic force. When the excited current gradually increased, it was able to overcome the weight of the movable axle center, which then moved upwards. The actuation of the EMA contained input, trip coil excitation, no-load motion of the trip coil axle center, axle center pushing trip trigger, main trip after tripping of trip trigger, and axle center returning to the home position after a trip action. The action process is detailed in Figure 2. The EMA actuation process looks simple, but its physical phenomenon is indeed complex. When the movable axle center moves upwards, according to Lenz\u2019s law, the trip coil generates an induced electromotive force, which not only suppresses the excited current, but also influences the magnetic force [12, 13]. This intercoupling continuously occurs inside the actuator. OF EMA The EMA actuation process involves three kinds of physical phenomena, including magnetic energy, electric energy, and mechanical energy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003170_icmid.2016.7738929-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003170_icmid.2016.7738929-Figure1-1.png",
        "caption": "Fig. 1. Process configuration of SLM process for development of complex 3D metal structures.",
        "texts": [
            " For functionalization, bronze alloy (CL 80CU with 90% Cu and 10% Sn) and pure copper in powder form and aluminum-oxide ceramic plates (thickness 0.65 mm) were used. The particle size diameter in the powder varies from approx. 9 \u03bcm to 50 \u03bcm. The main approach was to generate conductive patterns on the ceramic substrate through laser melting of the powder selectively to obtain highly dense and thick circuit patterns. In SLM process, the powder must be fully melted and simultaneously, melted up to a suitable depth for uniform adhesion and high density. The process follows basic RP technique as shown in Fig. 1, where the powder from feed chamber is applied in thin layers as build platform moves down. The predefined build process, part and material parameters control the platform heights, laser scan methodology and powder application. Numerous factors influence the bonding of the powder particles, which mainly involve laser velocity, laser scan speed, scan strategy, applied powder height as shown in Fig. 2. The melting of copper powder directly on ceramic surface and obtaining a reliable bond is a challenging task compared to melting on a metal platform"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002646_j.euromechsol.2016.07.002-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002646_j.euromechsol.2016.07.002-Figure2-1.png",
        "caption": "Fig. 2. (a) Scheme of a periodic bundle composed ofM piece-wise homogeneous fibers embedded in an elastic matrix. (b) Structure of a periodic cell composed of S segments of lengths l1, \u2026,lS (with PS n\u00bc1 ln \u00bc L). Each segment is described by a matrix A n containing all Young's moduli and interaction coefficients.",
        "texts": [
            " Accordingly to previous analysis for one fiber, the system is described by the following set of equations vti\u00f0x\u00de vx \u00bc Gi\u00f0x\u00de; (3) vui\u00f0x\u00de vx \u00bc 1 Ei\u00f0x\u00de ti\u00f0x\u00de; (4) for ci \u00bc 1;\u2026;M, where the force applied to each fiber is given by Gi\u00f0x\u00de \u00bc X j\u00bc1 M kij\u00f0x\u00de uj\u00f0x\u00de ui\u00f0x\u00de ; (5) describing the linear interaction among the fibers within the bundle. For the sake of generality, in this scheme we considered both the Young's moduli Ei and the interaction coefficients kij as functions of the abscissa x (heterogeneous system). In particular, this is useful to introduce a bundle structure characterized by a spatial period L. We suppose that each cell of periodicity is composed of S homogeneous bundle segments having lengths ln (n\u00bc1,\u2026,S) with PS n\u00bc1ln \u00bc L (see Fig. 2). Moreover, we consider all fibers with the same uniform cross-section S. For any homogeneous segment of the cell, we can write the elastostatic stressdisplacement interaction equations dti dx \u00bc X j\u00bc1 M k\u00f0n\u00deij uj ui ; (6) dui dx \u00bc 1 E\u00f0n\u00dei ti; (7) where ti(x) and ui(x) are longitudinal stress and displacement along the i-th fiber, k\u00f0n\u00deij are the interaction coefficients and E\u00f0n\u00dei represent the Young's moduli of the M fibers within the n th segment (i,j\u00bc1,\u2026,M and n\u00bc1,\u2026,S). It means that we have kij\u00f0x\u00de \u00bc k\u00f01\u00deij and Ei\u00f0x\u00de \u00bc E\u00f01\u00dei in the first segment (0<x<l1), kij\u00f0x\u00de \u00bc k\u00f02\u00deij and Ei\u00f0x\u00de \u00bc E\u00f02\u00dei in the second segment (l1<x<l1\u00fel2), and so forth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000619_j.proeng.2015.08.702-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000619_j.proeng.2015.08.702-Figure2-1.png",
        "caption": "Fig. 2: (a) Photo of a fabricated multi-parameter sensor chip (10 mm \u00d7 10 mm) including five individual working electrodes and (b) schematic sensor structure.",
        "texts": [
            " Glutaraldehyde was used for chemical cross-linking of L-LDH and DIA on the sensor surface. The sensor chip has been fabricated from a p-Si wafer by thin-film technology and photolithographic techniques. A detailed description of the fabrication process is presented in [10]. In the present work, the sensor was afterwards passivated with a 20 \u03bcm thick layer of SU-8 photo resist. The additional passivation provides an edge around the working electrodes, which facilitates the deposition of the liquid enzyme solution. In Fig. 2, a picture of the sensor chip, incorporated into a printed circuit board is presented. The corresponding sensor structure (cross section) is depicted schematically in Fig. 2(b). Before the enzyme immobilization, the sensor chip surface was cleaned with acetone, 2-propanpol and deionized water in sonication bath for 10 min each. The immobilization mixture was prepared by blending 2.4 \u03bcl of 10% bovine serum albumin (BSA) with 2.4 \u03bcl of enzyme solution (consisting of DIA and L-LDH dissolved in 0.1 M phosphate buffer, pH 7.6). Afterwards, 1.4 \u03bcl of a glutaraldehyde stock solution were added to the mixture. For varying the glutaraldehyde concentration (0.1-0.8%) in the enzyme membranes, different stock solutions were prepared"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002783_icphm.2016.7542869-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002783_icphm.2016.7542869-Figure1-1.png",
        "caption": "Figure 1: GE J85 engine and its accessory gearbox",
        "texts": [
            " On the other hand, the 1st GMF in the order spectrum is equivalent to the number of teeth of the target gear and the 2nd GMF is twice as that number, etc. Then the magnitude ratio is obtained from the spectrum by dividing the magnitude of the 2nd GMF by the magnitude of the 1st GMF. III. EXPERIMENTAL SETUP AND PROCEDURES The test facility used for this project was built at the National Research Council Canada (NRC). The test gearbox is an accessory gearbox of a J85 engine which is equipped on several aircrafts such as Canadair CT-114 Tutor. In real operation, the gearbox is integrated at the bottom of the engine, as shown in Fig. 1a and is mainly composed of four parallel gear shafts supported by rolling-element bearings (Fig. 1b). An overview of the test rig is shown in Fig. 2. The rig is reconstructed based on the prototype of the engine to obtain a practical operating environment. The rig is assembled on a steel frame and it is supported by a second steel frame which is firmly bolted to the ground. The upper and lower frames are spaced by rubber sheets with a purpose to isolate vibration transmissions from one to another. The gearbox is solidly mounted to the top bar of the upper frame. In order to simulate an operating environment of the test gearbox, the gearbox accessories were kept mounted on the various pads except for the fuel pump which was swapped by a truck alternator due to some technical issues with the fuel pump"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001214_ijhvs.2014.068101-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001214_ijhvs.2014.068101-Figure4-1.png",
        "caption": "Figure 4 Resistive action of creep forces (moment)",
        "texts": [
            " Referring to the axes system shown in Figure 3 considering origin at the centre of the elliptical contact patch and by using subscripts 3, 1 and 2 for forward, lateral and normal directions to the patch plane, the average velocity (Hobbs and Pearce, 1974), contact point velocity (Kalker, 1967) on the rail and the pure rolling velocity (Carter, 1926) of the wheelset have been used as denominator in equation (5) in different literature. Longitudinal creepage being very small this hardly makes any difference. Consequent upon the above discussion, the authors find it convenient to define creepage as: Corresponding slip velocityCreepage = . Pure rolling velocity (5) Here slip velocity means relative velocity of contact point on the wheel with respect to that on track. Dissipative (resistive) action of any component of creep force (moment) can be modelled by a modulated bond graph R-element (Figure 4) at 1-junction representing corresponding slip velocity. Using Kalker\u2019s linear theory: 3 33 33F f \u03be= (6) 1 11 11 12 22F f f\u03be \u03be= + (7) 2 12 11 22 22M f f\u03be \u03be= \u2212 + (8) where, F3 and F1 are creep forces in directions longitudinal and lateral to the patch plane and M2 is the creep moment. \u03be33 and \u03be11 are the creepages in the longitudinal and lateral directions and \u03be22 is spin creepage. fij are creep coefficients which are functions of wheelrail geometry, material properties and normal force on the patch plane. Considering equations (5)\u2013(8) and Figure 4 for contact \u2018a\u2019: 3 33 3.a a aF R S= (9) 1 11 1 12 2. .a a a a aF R S R S= + (10) 2 12 1 22 2. .a a a a aM R S R S= \u2212 + (11) In equations (9)\u2013(11) F3a, F1a and M2a represent creep forces in tangential and lateral directions and creep moment, respectively. Rija\u2019s are the values of corresponding R elements given by * 3/ .i ja af V * 3aV is pure rolling velocity at contact point \u2018a\u2019. 3 ,aS 1aS are the slip velocities in the forward and lateral directions and 2aS is the spin slip about the normal to the contact surface at \u2018a\u2019"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001115_s10894-015-9882-y-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001115_s10894-015-9882-y-Figure3-1.png",
        "caption": "Fig. 3 The benchmark of VV tolerance model a VV structure, b VV benchmark",
        "texts": [
            " The reference feature (benchmark) of the assembly should be defined firstly; then correlative dimension chains of every parts among assembly were added orderly, and error limits of every parts were determined according to actual process capacity; constrained relationships between parts were applied; geometric dimensioning allowance of every parts had been finished; check the tolerance analysis results till they could met assembly requirements. This can decrease the risk of calculation tolerance. Benchmark of Tolerance Model Dependent on assembly requirement, benchmark of the tolerance model involves three type of parts: VV (ribs), blocks and under bracket. Figure 3a is the VV structure, which supports for blocks and under brackets and it is basic component welded together with ribs. Figure 3b is the benchmark of VV tolerance model, in which three reference plane will be set up. DF-1 (front surface of VV front rib) is the reference plane of fourth block installing; DF-2 (central plane of right rib) is the reference plane of VV right rib surface (VV-R-F), which also is the standard of hole position on rib and it is perpendicular with DF-1; DF-3 (imaginary plane) is the perpendicular plane with DF-1 and DF-2 respectively. There are four blocks in the tolerance model. Although each has different shape and position, they have common point in installing, that is contacted with surface of under bracket firstly, then by means of upper bracket side surface and bolts are installed to holes of rib"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002624_jae-150151-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002624_jae-150151-Figure5-1.png",
        "caption": "Fig. 5. (a) Magnetic field lines for the generator of rated-load. (b) Electromagnetic torque for the generator of rated-load.",
        "texts": [
            " The exerted armature current can be got from the Eq. (17).\u23a7\u23a8 \u23a9 iA = Im sin(\u03c9t\u2212 \u03c3) iB = Im sin(\u03c9t\u2212 \u03c3 \u2212 120\u25e6) iC = Im sin(\u03c9t\u2212 \u03c3 + 120\u25e6) (17) By means of adjusting the excitation current and the successive iteration, the demand on the armature voltage U and the positive power factor cos\u03d5 should be satisfied as in the Eq. (18) and the Eq. (19). In this paper, \u03b5U is assigned as 0.2% and \u03b5cos\u03d5 is assigned as 0.5%. On the condition of rated-load, the magnetic field lines for the generator is shown in Fig. 5(a), and the electromagnetic torque for the generator is shown in Fig. 5(b).\u2223\u2223\u2223\u2223U \u2212 U\u03c6N U\u03c6N \u2223\u2223\u2223\u2223 \u03b5U (18) \u2223\u2223\u2223\u2223cos\u03d5\u2212 cos\u03d5N cos\u03d5N \u2223\u2223\u2223\u2223 \u03b5cos\u03d5 (19) To prove the validity of the dq0 positive and negative current vector algorithm for the second harmonic electromagnetic torque presented by this paper, the model of rated-load is revised by exerting negative current in the stator armature windings. Furthermore, when the negative current rate is the same, the initial negative current electrical angle is increasing to form arithmetic progression whose tolerance is \u03c0/6. The initial negative current electrical angle recorded by \u03b6r refers to phase angle difference between the instantaneous value of the negative current in phase-A and the instantaneous value of the positive current in phase-A"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002935_gt2016-56508-Figure18-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002935_gt2016-56508-Figure18-1.png",
        "caption": "Figure 18: Rotor assembly",
        "texts": [
            " The system may experience high vibration around 20krpm and 30 to 50krpm. The rotor-AFB system is highly stable up to 170krpm. At 190krpm to 200krpm, the backward whirling mode may be excited if there is external disturbance matched to the backward whirl frequency. Lastly, the vibration magnitude at the Aft bearing is predicted to be higher than at the Fwd bearing. Overall objective of the high speed test is to experimentally evaluate the rotor-AFB stability up to 185krpm. The rotor-AFB assembly is shown in Figure 18. To monitor the rotor vibration, four proximity sensors (P) are mounted in radial direction as shown in Figure 19. The Figure also indicates the location of the infrared high speed tachometer (S). Temperature around the bearing sleeves is monitored using four type K thermocouples (T). Cooling air (C) is supplied to the Aft AFB by feeding compressed air from the turbine shroud to the Aft bearing compartment. Figure 20 shows the overall test bed and instrumentation setup which also includes a computer where real time FFT signals can be monitored and a tachometer panel that displays the rotor speed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002064_2016-01-1136-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002064_2016-01-1136-Figure5-1.png",
        "caption": "Figure 5. Detail of the internal clutch\u2019s axial section.",
        "texts": [
            " The Archard\u2019s law [8], which is the most used method to analyze the wear trend in a contact between two components, defines the relation shown below (Eq. 2) among the wear depth \u03b4, the applied pressure p and the macroscopic displacement occurred during the contact s. Moreover, the relation considers two characteristics of the materials examined: the specific wear rate K and the hardness H of the softer material. (2) In this law, one of the most important assumption is the consideration of a constant pressure trend applied on the interfaces. In this analysis, due to the presence of a Belleville spring (fig. 5), the pressure on the discs\u2019 surfaces has a variable distribution. The contact patch between the spring and the first disc and between all the discs can be considered as circular crown areas, on which different radius dependent pressure distributions are built up. The axisymmetric geometry of the clutch allowed a two-dimensional finite element (FE) modelling of the assembly, with the advantage of obtaining a good trade-off between the processing speed and the simulation quality in terms of accuracy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001797_humanoids.2015.7363511-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001797_humanoids.2015.7363511-Figure9-1.png",
        "caption": "Fig. 9. Stepping motions against perturbations (120N, 0.5s) in diagonal direction.",
        "texts": [
            " Note that our stepping controller is not restricted to using the desired CoP indicators. Other indicators such as capture point or GFPE can be used as well. In simulation tests, we investigated the performance of the reactive stepping controller by applying perturbations in varying directions and magnitudes to the humanoid robot model (Figs. 8-10). Fig. 8 shows the different step lengths per different push strengths. When perturbed by a higher force, the robot naturally takes a longer step. Our controller can deal with perturbations in diagonal directions (Fig. 9). Fig. 9 (a) shows that our controller can take multiple steps if a single step cannot fully suppress the instability. The endurable range of magnitude of diagonal push is more limited than that of forward push because the discrepancy between the target stepping point and the actual stepping location increases due to the collision-avoiding maneuver of the swing foot explained in Sec. III-D. However, without the collision avoidance strategy, the swing foot easily bumps against the support leg, leading to the toppling of the robot as shown in Fig. 9 (b). (a) CoM trajectory (red curve) seen from the top (b) The trajectory of swing foot (right) in sagittal plane Fig. 11. CoM and swing foot trajectories during forward stepping. Fig. 11 shows the trajectories of CoM and the swing foot created by the stepping controller. It demonstrates that smooth curves emerge from our controller. The CoM trajectory is curved to the support foot to prevent toppling sideways, and it is similar to the trajectory of the well-known inverted pendulum model. The swing foot trajectory shows a smooth curve, but the overall shape is somewhat different from the hill-shaped curves used in other literatures [11], [12]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000032_b978-0-12-818482-0.00035-9-Figure35.11-1.png",
        "caption": "Fig. 35.11 Single and double row deep groove ball bearings.",
        "texts": [
            " The type of bearing indicated for each of the features should be considered as a first choice, but not necessarily the only choice. The bearings listed are described below. A e Deep groove ball bearing. B e Self-aligning ball bearing. C e Angular contact ball bearing. D e Cylindrical roller bearing. E e Needle roller bearing. F e Spherical roller bearing. G e Taper roller bearing. H e Thrust ball bearing. J e Spherical roller thrust bearing. K e Double row angular contact thrust bearing. 528 Manual of Engineering Drawing Deep groove ball bearings (Fig. 35.11) Deep groove ball bearings are available in both single row and double row designs. Single row ball bearings are the most popular of all rolling bearings. They are of simple design, non-separable and suitable for high-speed operation, and require little attention in service. The deep grooves enable axial loads to be carried in either direction. Bearings are available with shields and seals and can be supplied with the correct quantity of lithium base grease and used in operating temperatures between 30 C and \u00fe110 C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure2.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure2.5-1.png",
        "caption": "Fig. 2.5 (a) The deformation under shear stress. (b) Examples of loading conditions causing shear stresses between interfaces of glued blocks",
        "texts": [
            ", forming and extrusion) processes since they describe the material behavior during deformation. There are many different methods for mathematically representing the flow stress curves developed previously. The power law is a fairly good and simple approximation of the flow stress for materials during plastic deformation: \u03c3 \u00bc Ks\u03b5 n \u00f02:11\u00de where \u03c3 is the flow stress; Ks and n are constants; and \u03b5 is the strain during the material plastic deformation. 2.2 Solid Properties 39 Shear force is a force applied sideways on materials (transversely loaded) as described in Fig. 2.5a, not only in the solid state but also in the fluid state or the gas state. Shear stress, denoted by \u03c4, is the stress generated accordingly in the direction along the material surface and perpendicular to the tensile/compressive stress. Its unit is force per unit area carrying the load. The shear stress \u03c4 over a crosssectional area A can be generated when the material is being cut or pinned with a shear force F: \u03c4 \u00bc F A \u00f02:12\u00de Shear strain (\u03b3) is the ratio of the distance deformed to the height of the material, which can be measured by the angular change. The side lengths of the specimen remain unchanged, but the shape changes, like from a rectangle to a rhombus as shown in Fig. 2.5a. For a cubic material block exerted under a shear force with a side length L, a dislocation of the sheared surface has a dislocation of x. When the angle of side surface is very small, we can say that \u03b3 \u00bc x L \u00f02:13\u00de For the range of small deformations, the shear stress and the shear strain can be related to a material property called the modulus of rigidity (G): G \u00bc \u03c4 \u03b3 \u00f02:14\u00de We can conduct the mechanical test for determining the modulus of rigidity of a material with the known dimensions of a cross-sectional area A and the material thickness L by applying direct shear force (F) and simultaneously measuring the 40 2 Basic Material Properties surface displacement (x)",
            " This gradient can be converted to G by F x \u00bc L A \u03c4 \u03b3 \u00f02:15\u00de Like the elastic limit in the tensile stress-strain curve, if a material is sheared beyond a certain limit, yielding occurs, and the material deformation becomes irreversibly twisted and cannot return back to its original exact shape. If the material is stressed to the ultimate shear strength, the material will break. Shear stress can appear over interfaces between material bodies. To illustrate more details about interfacial shear stress, two cases of glued blocks under axial force are shown in Fig. 2.5b. In both cases, the blocks are in contact with other blocks by certain surface area a. In the left case, an external load F is acting on the side surfaces of the blocks, such that a shear force F is induced on the contact area. The right case contains three bodies with the side load of F/2 applied to the upper and lower blocks and the side load F in another direction. The load F on the middle block is distributed as two shearing forces over both the upper and lower interfaces, each with a shearing force of F/2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000015_intmag.2017.8007803-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000015_intmag.2017.8007803-Figure1-1.png",
        "caption": "Fig. 1 Spherical actuator and measurement equipment.",
        "texts": [
            " Due to this, it is difficult to measure the static torque characteristics. In this paper, in order to experimentally evaluate the static characteristics of a multi-DOF actuator, its flux linkage is firstly measured using a 3-RRR parallel mechanism [8], and its torque constant map is calculated. The static torque characteristics are evaluated using the torque constant map. Finally, the effectiveness of this method is discussed. II. Measurement method A. Torque Constant A 3-DOF spherical actuator in this study is shown in Fig. 1 (a)-(c) [7]. Spherical shell-shaped permanent magnets are arranged on the spherical mover, and coils are arranged on the stator teeth. Multiphase currents are applied to the coils to rotate the mover in 3-DOF. A torque equation is defined as (1), where K is a torque constant map, \u03c4X, \u03c4Y and \u03c4Z are current torques around X, Y and Z axes, respectively, and i1, i2, \u2026 and in are coil currents of each phase. In order to experimentally evaluate the static characteristics of the actuator, the torque constant map of a real actuator is created",
            " The number of components that a correlation coefficient of the original value and interpolation value is 0.999 or more were obtained. The measurement points that reduce errors due to data interpolations are obtained by Genetic Algorism (GA). This process will be described in the full paper. C. Measurement Equipment Unlike 1-DOF actuators, a spherical actuator cannot rotate continuously in arbitrary axis, and it is difficult to measure the induced voltage. Therefore, the torque constants are calculated from the measured flux linkage. In order to rotate the mover, a 3-RRR parallel mechanism shown in Fig. 1 (d) is used. This mechanism is composed of an end effector, links and 3 servo motors. It is possible to rotate the actuator in 3-DOF using 9 rotational axes. By using inverse kinematics, the position of the end effector is determined. The torque constants are calculated using (2), where \u03c6 is a flux linkage of the single pole, \u0394\u03b8 is a small displacement. III. Results The comparison of measured and analyzed torque constants are shown in Fig. 2 (a). The horizontal axis shows the measurement position decided by GA"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001510_s0263574713001203-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001510_s0263574713001203-Figure4-1.png",
        "caption": "Fig. 4. Input control voltages (proposed controller).",
        "texts": [
            "126 Figure 3 shows the end-effector tracking errors for the proposed controller. The solid line shows the end-effector tracking error along x axis and the dashed line shows the end-effector tracking error along y axis. Looking deeply into Fig. 3, we see that the robot cannot track the desired trajectory in the first 4 seconds, but it can estimate the parameters and variables after that and the performance is satisfactory after 4 seconds. Input control voltages for the first and second joints are shown in Fig. 4. The solid line shows the input control voltage for the first joint and the dashed line shows the input control voltage for the second joint. From Fig. 4, we see that the proposed controller introduces chattering into the robot system. To eliminate the chattering effects, the signum function sgn( \u02dc\u0307q) can be replaced by tanh(\u03b5 \u02dc\u0307q). Input control voltages of applying the proposed controller with tanh(\u03b5 \u02dc\u0307q), \u03b5 = 10 are shown in Fig. 5. It can be seen that the chattering is completely eliminated. In order to compare the results of the proposed controller with the previous works, we apply the passivity-based controller in ref. [20] on the robot. Initial conditions are the same as the proposed http://journals"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001506_s11094-014-1146-z-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001506_s11094-014-1146-z-Figure1-1.png",
        "caption": "Fig. 1. Schematic of flow-through three-electrode cells: with round electrodes (a), with planar electrodes (b ); reference electrode (1 ), working electrode (2 ), auxiliary electrode (3 ).",
        "texts": [
            " The composite IrO x \u2014RuO x was immobilized by sequential electrochemical deposition of each metal oxide. The deposition occurred from solutions containing RuCl 3 (pure, Aldrich) and IrCl 3 (pure, Acros Organics) in KNO 3 (0.1 M) and H 2 C 2 O 4 (0.05 M), respectively. The solutions were prepared by dissolving accurately weighed reagents. FIA was performed on the previously described appara- tus [12]. We used two types of electrochemical flow-through cells. The first had three round electrodes and cell volume 200 L (Fig. 1a ). The second was a wall-jet cell with planar electrodes and volume 40 L (Fig. 1b ). We used allopurinol (pure, Aldrich) with the following structural formula: The allopurinol solution was prepared by dissolving an accurately weighed portion in aqueous supporting electro- a b 6.86) (a); plot of I vs. for allopurinol oxidation on a CME with IrO x \u2014RuO x \u2014SWCNT composite (b ). N N O N H NH lyte. A series of solutions of decreasing concentrations were prepared by diluting stock solutions immediately before use. Drug solutions were also prepared by dissolving accurately weighed portions in supporting electrolyte",
            " This was due to the formation of a highly dispersed nanostructured deposit on the electrode sur- face, in contrast with the other CME [17]. It is noteworthy that the catalytic response of the CME with IrO x \u2014RuO x \u2014SWCNT was highly stable and reproducible. A method for amperometric detection of allopurinol at a CME with IrO x \u2014RuO x \u2014SWCNT under FIA conditions was developed. Various types of flow cells and working electrodes were used for the research. These included a three-electrode cell with round electrodes and a wall-jet cell with planar electrodes (Fig. 1). The dependence of the CME catalytic response in the flow-through system on electrochemical parameters such as applied potential (E) and hydrodynamic parameters such as sample injected volume (V) and flow rate (u) was studied. The FIA conditions were determined based on the obtained experimental results. Thus, the analytical signal in the flow-through cell with round electrodes was recorded at E = 1.25 V, V = 0.60 mL, and u = 15.5 mL/min; in the flow-through cell with planar electrodes, at E = 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003621_b978-0-12-382038-9.00008-9-Figure8.10-1.png",
        "caption": "FIGURE 8.10",
        "texts": [
            " The contact between the two members of a higher pair joint has point or line geometry. The contact stress for a higher pair joint is large because the contact area is very small. If there is pure rolling contact between the members, then there is no relative sliding between the contact surfaces and thus friction and wear are minimal. The number of degrees of freedom for a higher pair joint can be large as the point or line contact allows for less constrained motion of members. A cam-follower, as shown in Figure 8.10, is a good example of higher pair joints, where a line contact is observed between the cam and the follower. A cam-follower allows two DOF, one rotational and one translational, along the center axes of the cam and the followers that are in parallel. Lower pair and/or higher pair joints are combined as per the design requirement to obtain compound joints. Compound joints composed of higher pair joints can be kinematically equivalent to lower pair joints and vice versa. By such combinations desirable features from the combining joints are retained to achieve robust joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000072_978-1-4614-9126-2-Figure7.54-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000072_978-1-4614-9126-2-Figure7.54-1.png",
        "caption": "Fig. 7.54 Multi-effect evaporator",
        "texts": [
            " In order to design an industrial process to evaporate much of the water present in the product X4, an experiment was conducted in a laboratory-scale evaporator (simple) as shown in Fig. 7.53: It was experimentally shown that this evaporator is capable of evaporating 80 % of the water initially contained in the product. The X4 product contains 80 % w/w water, and it is expected that the final product will not be more than 10 % w/w water and no less than 6 % w/w. As the water content of the product obtained in the laboratory equipment is greater than 10 %, in practice, it will be necessary to build an evaporator as depicted in Fig. 7.54: From experience, we know that each evaporator added to the process is less efficient than the preceding one. The percentage of water that is removed at each stage is determined by the formula X \u00bc 90 10i, where X is the percentage of H2O removed in stage i and i is the stage (1,2,3,. . ., N). (a) How many evaporators (stages) do there have to be so that the product meets the requirements established for the X4 end product? (b) What is the final concentration of H2O in the end product? A: (a) 3 (b) 8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000790_eml.2014.6920669-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000790_eml.2014.6920669-Figure8-1.png",
        "caption": "Fig 8 Eddy current density distribution on load",
        "texts": [
            " Fig 6 gives out the harmonic order on the surface of magnet due to static eccentricity. Compared with noneccentric rotor, the main harmonic order is 1/2and 3/2, which agrees with the equation (10). Table 1 gives the rotor loss when the static eccentricity is 0.6. Both magnet loss and rotor yoke loss are more than non- eccentric rotor. The total loss increases nearly 30 times, but the loss on each magnet increases almost the same. The eddy current density distribution is compared to a noneccentric rotor in Fig 7and Fig 8. It has been seen that with 0.6eccentricity, the eddy current density has significant changes. According to the equation (22), it is obvious that the eddy current loss on magnet is proportion to the square of eccentricity . Fig 9 shows that the loss increases with the increase of eccentricity. From the above analysis, it is seen that static eccentricity has a visible effect on magnet loss. Hence we should try to avoid large static eccentricity and make better efficiency and lower magnet loss. B"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001599_ijamechs.2014.066928-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001599_ijamechs.2014.066928-Figure14-1.png",
        "caption": "Figure 14 Escape from deadlock (see online version for colours)",
        "texts": [
            " In addition, the size of the duplex system is reduced from that of the conventional one. The locking mechanism was also improved to realise the semi-circular shape in Figure 12. By pushing the piston attached to the end of the manipulator, the hose is expanded by air, and the pins engage to the holes of the link. This locks the joint. To solve P2, a pulling mechanism was employed. Figure 13 shows the pulling mechanism. Both manipulators have the same mechanism, but only one is shown in order to simplify the illustration. As shown in Figure 14, by pulling the wire of the locked manipulator, the movable manipulator is pulled forward. This reduces the friction, and the duplex manipulator can escape from its deadlock. By this mechanism, we can utilise both pushing and pulling operations. The pushing operation is effective for the outside manipulator, and the pulling operation is effective for the inside manipulator. The mechanisms for turning the head link are similar to those of our previous manipulator described in Section 3. Details on the developed manipulator are presented in Section 6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000617_ascc.2015.7244379-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000617_ascc.2015.7244379-Figure1-1.png",
        "caption": "Fig. 1. Generalised Stewart platform manipulator.",
        "texts": [
            " Similarly the vector Bi which describes the position of the base attachment points with respect to frame Fb is Bi = [Bix Biy Biz] T Then the six degree of motion state including three axis linear translations and three axis rotations is represented as X = [px py pz \u03b1 \u03b2 \u03b3 ]T (1) where px, py and pz represent the surge, sway and heave motions respectively and \u03b1, \u03b2 and \u03b3 represents the roll, pitch and yaw motions of the platform with respect to the three axes of the fixed frame. Each ith leg links the attachment points Pi of the platform and Bi of the base plate respectively as shown in Fig.1. Generally the inverse kinematics of the parallel manipulator has a unique solution. Here for a given position and orientation of the platform corresponding leg length can be uniquely determined by inverse kinematic formulation. Then the leg vector pointing from attachment point Bi to attachment point Pi can be expressed as follows li = R\u03b1\u03b2\u03b3Pi +D\u2212Bi (2) where the matrix R\u03b1\u03b2\u03b3 represent the orientation of the frame Fp with respect to frame Fb being expressed as R\u03b1\u03b2\u03b3 = [ c\u03b1c\u03b2 c\u03b1s\u03b2s\u03b3 \u2212 s\u03b1c\u03b3 c\u03b1s\u03b2c\u03b3 + s\u03b1s\u03b3 s\u03b1c\u03b2 s\u03b1s\u03b2s\u03b3 + c\u03b1c\u03b3 s\u03b1s\u03b2c\u03b3 \u2212 c\u03b1s\u03b3 s\u03b2 c\u03b2s\u03b3 c\u03b2c\u03b3 ] (3) here c(\u00b7) = cos(\u00b7) and s(\u00b7) = sin(\u00b7)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003271_978-3-319-30897-5_4-Figure4.25-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003271_978-3-319-30897-5_4-Figure4.25-1.png",
        "caption": "Fig. 4.25 Spring-damper-actuator element connecting bodies i and j",
        "texts": [
            " The choice of the negative y direction as the direction of gravity is totally arbitrary. However, in the present work, the gravitational field will be considered to be acting in this direction unless indicated otherwise. The force due to gravitational field can be written as FG \u00bc mig~uy \u00f04:93\u00de where mi is the mass of body i and g is gravity acceleration. In the present work the acceleration due to gravity assumes a value equal to 9.81 m/s2 in the SI system of units. The module of the gravitational force is given by FG \u00bc mig \u00f04:94\u00de Figure 4.25 shows a spring-damper-actuator element connecting bodies i and j through two points of connectivity Pi and Pj, whose local position vectors with respect to the origins of their respective local coordinate systems are s0Pi and s0Pj . The spring-damper-actuator force is given by FSDA \u00bc k\u00f0l l0\u00de\u00fe d_l\u00feFA u \u00f04:95\u00de where the first term on the right-hand side is the spring force, the second term represents the damper force and the third term denotes the actuator force. The spring stiffness is represented by k, l is the deformed length, l0 is the undeformed or natural length of the spring, d is the damping coefficient of the damper and _l is the time rate of change of the damper length"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002933_gt2016-56900-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002933_gt2016-56900-Figure5-1.png",
        "caption": "Figure 5 B-H curve of of the AMB-HPS laminate material",
        "texts": [
            " Figure 4 shows the basic structure of silicon iron sheets that consist the AMB stator. Its dimensional parameters are listed in Tab.1. The rotor and stator of the acturator are laminated to reduce the eddy current losses in the integrated AMB-HPS. The Silicon iron is selected as the lamination material for both the rotor and the stator for its low magnetic losses and relatively low cost compared to those of other magnetic materials. The type of material used in the lamination of the rotor and stator is 0.35 mm gaute. The B-H curve of the material is indicated in Fig.5 Table 1 Parameters of electromagnetic actuator Type Item Value Basical parameters Number of turns in N poles 100 Number of turns in S poles 75 Diameter of rotor 125 mm Diameter of stator 126.2mm Rotor-stator clearance (g0) 0.6 mm Width of rotor and stator (l) 60 mm Large pole angle (\u03b31) 22.85\u00b0 Small pole angle (\u03b32) 11.37\u00b0 Large pole opening angle (\u03c31) 15.89\u00b0 Small pole opening angle (\u03c32) 12.63\u00b0 Pole location angle (\u03b2) 33\u00b0 Hole parameters Hole diameter (D) Hole depth (l1) Hole distance (d) Number of holes (n) 3 mm 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003559_s1474-6670(17)70080-8-FigureI-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003559_s1474-6670(17)70080-8-FigureI-1.png",
        "caption": "Figure I I. Optimum coefficients kl and k2 for systems with varying l' and fixed D",
        "texts": [
            " Owing to imperfections, this means reaching a stable limit cycle around y == Xo and y' = y\" == O. (c) The input is a continuous function of time. In this case, best performance usually means that the error should be a minimum at all times after initial disturbances have been overcome. Occasionally, if the input is a continuous function which either is a short time disturbance or represents practically a step input, statements (a) and (b) may suggest better solutions. For the general continuous input the output should approxi mate the input in a serpentine fashion (see Figure I). This means that the output should have a very high frequency compared to any frequency contained in the input. The System Under Consideration The dynamic system to be controlled is given by the equation y''' + (2D + y)y\" + (1 + 2Dy )y' ,;+- YI' =: :X2x + :XIX + :Xll'\\: (l) and is represented in Figure 2*. The prime represents differenti ation with respect to the dimensionless time T. The parameters 0(0' 0(1) 0(2' Y and D are real numbers. If D > 1 the transfer function has only real poles. This case has been frequently treatedl - 3 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000540_1754337115582121-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000540_1754337115582121-Figure1-1.png",
        "caption": "Figure 1. Schematic of the oar-shaft\u2019s dynamic behavior during the drive. The equilibrium position (Ex) is the point where the magnitude of the oar-shaft\u2019s deflection in the x-axis is zero. d is deflection and d21 is inverse deflection, as described in the text. Deflection of the inboard is neglected in this model.",
        "texts": [
            "owing, oar-shaft, biomechanics, stiffness, deflection, lever Date received: 5 September 2014; accepted: 24 March 2015 The effect of oar-shaft stiffness on rowing biomechanics is not well known. Many previous studies have assumed that the oar-shaft is perfectly rigid.1\u201311 The dynamic behavior of the oar-shaft during the drive is illustrated schematically in Figure 1. The equilibrium position (Ex) is the point where the magnitude of the oar-shaft\u2019s deflection is zero (i.e. during the recovery when there is no load on the blades\u2014neglecting air resistance). Following the catch position, the blades enter the water and the rower pulls on the handles. The oar-shafts deflect (d) toward the bow as the blades experience resistance while moving through the water; this deflection stores elastic potential energy in the shaft\u2019s material. Toward the end of the drive, the rower\u2019s force application to the handles decreases and the oar-shafts inversely deflect (d21) back to their Ex position"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001745_omae2015-41955-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001745_omae2015-41955-Figure3-1.png",
        "caption": "Figure 3. CROSS SECTION OF A HYSTERESIS IPM MOTOR.",
        "texts": [
            " HYSTERESIS IPM MOTOR A hysteresis IPM motor\u2019s rotor has a cylindrical ring made of composite material like 17% or 36% cobalt steel alloy, special Al-Ni-Co, Vicalloy, etc. with high degree of hysteresis energy per unit volume [6, 11]. The rare earth permanent magnets are buried inside the hysteresis ring and the ring is supported by a sleeve made of non-magnetic materials like aluminum which forces the flux to flow circumferentially inside the rotor ring [6, 11]. Fig. 2 illustrates the rotor of a hysteresis IPM motor. The cross section of a hysteresis IPM motor depicting the position and orientation of permanent magnets is shown in Fig. 3. The inclusion of permanent magnets creates rotor saliency without changing the length of the physical airgap and provides an additional permanent source of excitation in the rotor. Magnetic hysteresis refers to the dependency of a material\u2019s magnetization on the past states of the material, as well as its current state. Fig. 4 shows the major hysteresis loop of a hysteresis ring made of 36% Cobalt-Steel alloy. Br is called the remanence which is the residual value of flux density when the applied field becomes zero"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001298_j.finel.2014.06.011-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001298_j.finel.2014.06.011-Figure1-1.png",
        "caption": "Fig. 1. Satellite tracking antenna.",
        "texts": [
            " If these effects cannot be neglected, or for instance very precise predictions of the eigenfrequencies are vital, transforming the system matrices into a 2n or 3n-space provides a way of extracting the more correct eigenfrequencies, however, this comes at the cost of an increase in solving time. It should be noted that the transformation of the system matrices into a 2n or 3n-space does not necessarily mean an impractical or unsolvable equation system with respect to number of system DOFs. An example given by the authors in [3] is the satellite tracking antenna depicted in Fig. 1. Due to model reduction techniques [2], the virtual model of the mechanism was reduced from approximately 950,000 DOFs to about 850 DOFs for effective time domain dynamic simulations. The objective of this work is to help engineers working in an FE environment to be able to accurately predict eigenfrequencies and mode shapes of active mechanisms containing any type of PID controllers, with the exception being controllers containing acceleration feedback derivative gains. The controllers can be of type single-input single-output (SISO) or multiple-input multiple-output (MIMO), and the sensors and actuators for the controllers can be either collocated or non-collocated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000816_s10409-015-0016-6-Figure7-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000816_s10409-015-0016-6-Figure7-1.png",
        "caption": "Fig. 7 Contours of the dimensionless axial velocity on a the horizontal plane \u03c2 = z/a = 1.0 and b the vertical plane \u03b7 = y/a = 0",
        "texts": [
            " In contrast to the dimensionless pressure, which is symmetric with respect to the horizontal plane z = 0, the dimensionless radial velocity is antisymmetric with respect to the same plane, as shown in Fig. 5b. Figure 6 plots the dimensionless axial velocity as functions of the dimensionless coordinates. It is seen from Fig. 6 that V\u03c2 generally deceases with both \u03c1\u2032 = \u03c1/a and \u03c2 = z/a. In particular, for \u03c1\u2032 > 1.0, V\u03c2 increases slightly with \u03c2 = z/a in the neighborhood of \u03c2 = 1.0. The velocity with a radial coordinate \u03c1\u2032 1.0 is much larger than that corresponding to \u03c1\u2032 > 1.0. For completeness, the contours for V\u03c2 are shown in Fig. 7. As expected, the contours on the horizontal plane \u03c2 = z/a = 1.0 are a series of concentric circles (Fig. 7a). In contrast to V\u03c1\u2032 and p\u2032, V\u03c2 is neither symmetric nor antisymmetric with respect to the plane \u03b7 = 0 as shown in Fig. 7b. Due to the axial translation in the z-direction, the axial velocity V\u03c2 0 means that the fluid flows in the same direction as that of the plate. Further, the fluid near the plate possesses a larger axial velocity. These observations are in agreement with physical intuitions. 6.2 Elliptic plate Figure 8 plots the dimensionless force P\u2032, required to be exerted on the elliptic plate, as a function of the eccentricity e. It is seen that P\u2032 changes significantly with e. As expected, P\u2032|e=0 is equal to 16, which corresponds to that for a circular plate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003816_978-981-10-0733-0-FigureB.2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003816_978-981-10-0733-0-FigureB.2-1.png",
        "caption": "Fig. B.2 Configuration for the parallel-axis theorem",
        "texts": [
            "4) 4A better expression would be moment of surface since area means strictly speaking the measure of the size of the surface which is different to the surface itself. \u00a9 Springer Science+Business Media Singapore 2016 A. \u00d6chsner, Computational Statics and Dynamics, DOI 10.1007/978-981-10-0733-0 407 408 Appendix B: Mechanics Fig. B.1 Plane surface with centroid S B.2 Parallel-Axis Theorem The parallel-axis theorem gives the relationship between the secondmoment of area5 with respect to a centroidal axis (z1, y1) and the second moment of area with respect to any parallel axis6 (z, y). For the rectangular shown in Fig.B.2, the relations can be expressed as: 5The second moment of area is also called in the literature the second moment of inertia. However, the expression moment of inertia is in the context of properties of surfaces misleading since no mass or movement is involved. 6This arbitrary axis can be for example the axis trough the common centroid S of a composed surface. Appendix C Units and Conversion C.1 SI Base Units The International System of Units (SI)7 must be used in scientific publications to express physical units"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003558_b978-0-12-801578-0.00012-6-Figure12.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003558_b978-0-12-801578-0.00012-6-Figure12.9-1.png",
        "caption": "Figure 12.9 (Left) Photoinduced rolling motion of continuous ring of LCE film induced by simultaneous irradiation with UV (366 nm, 200 mW/cm2) and visible light (>500 nm, 120 mW/cm2) at room temperature. Size of LCE ring 18 mm 3 mm 20 mm; (center) rotation of light-driven motor using LCE laminated plastic film: (right) schematic illustration of the motor. Reprinted with permission from Yamada, M.; Kondo, M.; Mamiya, J. I.; Yu, Y.; Kinoshita, M.; Barrett, C. J.; Ikeda, T. Photomobile Polymer Materials: Toward Light-Driven Plastic Motors. Angew. Chem. Int. Ed. 2008, 47, 4986e4988. Copyright 2008 Wiley and Sons.",
        "texts": [
            "66 prepared a continuous ring of such an LCeelastomer (LCE) film using polymeric azobenzene meoities (a nice review of photoresponsive polymeric micelles has recently been written by Huang et al.67). Upon exposure to UV light in specific locations, the ring rolled toward the light source, almost completing a full revolution at room temperature. This ring was mounted to a pulley system, which, when irradiated simultaneously with UV and visible light, caused both local contraction and expansion forces. This combination gave rise to the observed photoinduce motion and the driving of the pulleys (Fig. 12.9). Furthering this work, and by employing an LC polymer network that can selectively form either right- or left-handed macroscopic helices, Iamsaard and coworkers demonstrated a nanoscale energy convertor whose photoinduced deformations scaled with the concentration of the azobenzene and could achieve useful work. This process the authors compared to the movement of plant tendrils.68 Light-responsive WLMs take advantage of the light-induced dimerization,69 cisetrans isomerization of surfactants,70 or additives containing a suitable chromophore"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002773_9781782421955.1049-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002773_9781782421955.1049-Figure1-1.png",
        "caption": "Figure 1 \u2013 Top: Physical installation. Bottom: Three-dimensional representation of the test rig: 1. Test gears; 2. Reaction Gears; 3. Bearings support plates; 4. Flexible couplings; 5. Flywheels; 6. Clutch flange for preload. More detailed analysis on the instrumentation and the design of the test rig can be found in (1). The main specifications of the rig are reported in Table 1, with a special highlight for precise introduction of parallel and angular misalignments imposed using the method described in (2).",
        "texts": [
            " Modal analysis has been performed on the test rig first at a component level in free-free conditions, then at a sub-assembly level and finally at the full assembly level. The effects of gear meshing, bearings and couplings have been evaluated for mode shapes and natural frequencies. Although there is substantial structural coupling between the gears and the shaft bending modes, it is observed that the dynamic response of the gear pair is dominated by the first torsional mode due to the elastic deflection of gear teeth. 1049 2 TEST RIG DESCRIPTION The main aim of the test rig, shown in Figure 1, is to allow the heavy instrumentation of a test gear pair which can be subject to tightly controlled operating conditions and which has well-characterised boundary conditions. Parallel misalignments 0 to 0.3 mm 0.020 mm Instrumentation includes triaxial accelerometers mounted on the gear body to measure dynamic relative 5-DoF displacements (3); two high resolution analogue encoders (18000 sinusoidal periods per rotation) and two lower resolution digital encoders (5000 pulses per rotation) mounted at the shaft ends to measure transmission error; a torque sensor and strain gauge arrays mounted at tooth roots"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure2.8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure2.8-1.png",
        "caption": "Fig. 2.8 (a) Configuration and (b) velocity profile of the Couette flow",
        "texts": [
            " Fluids with high viscosity are \u201choneylike\u201d \u2013 their flow is very slow and resistive since the high internal friction between fluidic molecules resists motion. Viscosity is important for evaluating the quantity of fluids in transportation during a certain period of time and the energy losses due to transport of fluids in tubing, syringes, channels, and slits in manufacturing processes and even for vessels in the human body. In order to examine the role of viscosity, we may first consider the Couette flow. It is a shear force induced laminar flow of fluid between two adjacent plates with the upper plate moving along one direction (Fig. 2.8a). If we assume a consistent viscosity \u03bc over the fluid volume with a thickness H and impose a constant velocity of the moving plate U, we will have a consistent shear stress \u03c4 generated in the fluid volume between the plates, and the velocity profile inside the fluid u( y) varying with 2.4 Fluidic Properties 45 the horizontal position y can be described as Fig. 2.8b. The role of viscosity in the associated shear stress can be explained via the flow between the two no-slip parallel plates as \u03c4 \u00bc \u03bc du dy \u00f02:28\u00de We may further consider the no-slip condition where the boundaries of fluid adhering to the plates show zero velocity relative to the plate walls on which the fluidic adhesion force to the wall is stronger than the cohesion force by flow. Hence, we obtain u(y \u00bc 0) \u00bc 0 and u(y \u00bc H ) \u00bc U. Therefore, u( y) \u00bc Uy/H and \u03c4 \u00bc \u03bcU/H. Most common fluids, like water, air, and oil, exhibit a linear relation between shear stress (\u03c4) and shear rate (du/dy) by the factor of viscosity \u03bc"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002247_s12204-014-1477-7-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002247_s12204-014-1477-7-Figure5-1.png",
        "caption": "Fig. 5 Coordinate system",
        "texts": [
            " Conversely, if the coordinates of the torch are known, and the working and moving angles are given, the optimal posture of the manipulator can also be obtained. However, the calculation process is too cumbersome. In the following discussion, a calculation method in the condition of special moving angle will be introduced. A variety of manipulator postures may lead to the fact that the torch can reach the same point. With a certain working angle and a certain moving angle, the posture is fixed. In the discussion, special moving angle is \u03a8 = 0\u25e6, which means that the welding torch axis must be vertical to the welding direction. As shown in Fig. 5, the world coordinate system is established. In the coordinate system, the vehicle forwarding direction is x axis. In the plane the direction which is perpendicular to x axis is y axis. And the direction which is perpendicular to the plane is z axis. The angle between manipulator and y axis is the joint angle of \u03b81. Joint angles of \u03b82, \u03b83 and \u03b84 determine the extended length of manipulator, while the joint angle of \u03b85 determines the moving angle of the torch. In the figure, we set the coordinates of point P as Pi(xi, yi, zi)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001556_1.4931346-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001556_1.4931346-Figure1-1.png",
        "caption": "FIGURE 1. A 2-D Closed-Kinematic Biomechanical Model of Ten Segments",
        "texts": [
            " Several studies on human models have been done quite extensively such as modelling of lower limb muscles during human maximaleffort countermovement jumping [13], determination of torques at upper limb joints during jumping in badminton smash via Kane\u2019s Method [14], musculoskeletal model for backhand frisbee throws [15], modeling of an arm via Kane\u2019s method [16], modelling of lower limb using Kane\u2019s method from a jumping smash activity [17] and a sixlink kinematic chain model of human body using Kanes\u2019s method [18]. However, as far as today, a full human model of a harvester during harvesting Fresh Fruit Bunch (FFB) via Kane\u2019s method has never been reported. So, the main purpose of this article is to develop a mathematical model of human body during harvesting activity via Kane\u2019s method. MODEL A two-dimensional closed-kinematic biomechanical model representing a harvester movement is developed using Kane\u2019s method (Figure 1). The ten-segment model consisted of the foot, the leg, the trunk, the head and the arm segments. Kane\u2019s method is a vector-based approach which uses vector cross and dot products to determine The 2015 UKM FST Postgraduate Colloquium AIP Conf. Proc. 1678, 060019-1\u2013060019-8; doi: 10.1063/1.4931346 \u00a9 2015 AIP Publishing LLC 978-0-7354-1325-2/$30.00 060019-1 velocities and acceleration rather than calculus [19]. It creates auxiliary quantities called partial angular velocities and partial velocities, and uses them to form dot product with the forces and torques acting from external and inertial forces"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000422_978-94-007-4620-6_50-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000422_978-94-007-4620-6_50-Figure1-1.png",
        "caption": "Fig. 1 General hexagonal Stewart platform manipulator (GHSPM).",
        "texts": [
            " The most general form of SPM consists of a rigid moving platform connected to a fixed rigid base by six identical legs, of UPS or SPS architecture. However, in practice, the fixed as well as the moving platforms are of the form of rigid hexagons1 in most SPMs. This class of manipulators, 1 The case of triangular platforms is obviously included in this group. known as the general hexagonal Stewart platform manipulator (abbreviated here as GHSPM), is studied in this paper. The manipulator is shown in Fig. 1. The legs are assumed to have UPS architecture (which has the same kinematics as the SPS-legged SPM, except for the idle rotations of the legs about their respective axes). Let t i = (xti , yti , 0)T denote the position of the ith S-joint in the top platform with respect to the coordinate system attached to a point p(x, y, z) on the platform. Similarly, let bi = (xbi, ybi, 0)T denote the corresponding points in the base coordinate system. The rotation matrix R, expressed in terms of the Rodrigue\u2019s parameters c1, c2, c3, represent the orientation of the moving frame with respect to the fixed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003036_978-3-319-46669-9_199-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003036_978-3-319-46669-9_199-Figure1-1.png",
        "caption": "Fig. 1 Robotic walker",
        "texts": [
            " Compared with over-ground walking, walking on a treadmill indicates significantly greater cadence, smaller stride length, stride time and reductions in joint angles, moments, powers and pelvic rotation excursion [6]. Thus, for enhancing gait performance and presenting natural gait patterns with actual foot contact, rehabilitation devices with over-ground walking are more recommendable. A more effective robotic system with sufficient assistance for pelvic motions is critical to overcome limitations in current devices and to bring benefits to a broad spectrum of patients. In this paper, we introduced the design and experimental results of a novel robotic walker (shown in Fig. 1a) with pelvic motion assistance and over-ground walking. As shown in Fig. 1a, the robot consists of an omni-directional mobile platform integrated with an active body weight support unit and wearable modular body motion sensors. It has 4 active and 2 passive DOFs to perform pelvis movements, shown as in Fig. 1b. The omni-directional mobile platform supports pelvic motions in Anterior-Posterior (AP, Vy), Medio-Lateral (ML, Vx), and rotational (RT, \u03c9) movements. These three DoFs of pelvic motions are implemented by two sets of ASOC (active split offset castor) units, consisting of two coaxial conventional wheels with All-in-One Hub Motors. The pelvic brace consists of a harness and pads to passively support both pelvic tilt and obliquity. By applying the pelvic support brace, unloading certain percentage of the body weight is achieved to support patients with weak muscles to practice gait training more in this design"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001399_1464419315571983-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001399_1464419315571983-Figure5-1.png",
        "caption": "Figure 5. Pulse excitation caused by the localized defect. (a) Pulse generation. (b) Three different pulse forms.",
        "texts": [
            " Considering the additional dynamic load caused by the bearing vibration, then these excitation loads can be calculated using Fvcx t\u00f0 \u00de \u00bc Ktx Ktx Fx \u00fe ~Fx Fvcy t\u00f0 \u00de \u00bc Kty Kty Fy \u00fe ~Fy Mvcx t\u00f0 \u00de \u00bc K tx K tx Mx \u00fe ~Mx Mvcy t\u00f0 \u00de \u00bc K ty K ty My \u00fe ~My 8>>>>>>>>>>>< >>>>>>>>>>>: \u00f015\u00de Taking the additional dynamic load (the frequency is !s) due to the unbalance response into consideration, the VC excitations can be expressed as the summations of a series of harmonic components at different frequencies in which the VC frequency, !vc, the rotational frequency, !s, and the sums and the differences of these frequencies are included. As shown in Figure 5, when a rolling element rolls over a small localized race defect in the loaded zone of the bearing, the contact load will reduce in comparison to that of the rolling element-healthy race contact and an impact excitation will form if the static displacement of the bearing remains unchanged. Because of its short duration, some researchers16,22 tended to consider the impact excitation due to the localized defect as an impulse without duration. However, the severity, extent, and age of the localized defect can be better represented by the pulse with finite width",
            " For a constant rotational speed, the pulse generated due to the localized defect is periodic in nature, and the frequency of the pulse generation depends on the location of the defect. The width, T, of the pulse form, f \u00f0t\u00de, can be determined by dividing the defect width by the relative velocity between the rolling element and the race. The pulse form depends on the severity, extent, and age of the damage and it may not be of a very regular shape. Instead of the actual pulse form, rectangular pulse, triangular pulse, and half-sine pulse as shown in Figure 5 are usually employed in many studies to simplify the problem. In the practical application of water pump bearings, localized defects are always generated on the bearing races and almost no localized defects occur on rolling elements, so only the localized defects on inner and outer races are considered here. Rolling elements of the bearing are assumed to be massless elastic bodies in this paper because the mass of a rolling element is much smaller than those of the at Monash University on September 18, 2015pik"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003991_t-aiee.1935.5057113-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003991_t-aiee.1935.5057113-Figure4-1.png",
        "caption": "Fig. 4. Visualization of cap surface for a dynamo close approximation to a rectangle, and the field is a",
        "texts": [
            " This valued function, the magnetic circuit must be capped component gives the usual formulas for pole core with a diaphragm which may serve as a datum from leakage, although most books underestimate the which to reckon magnetomotive force. In general leakage from the pole ends.6 The compensating this surface is arbitrary, but here considerations of field will give a new term to be considered. symmetry almost require that the neutral axis be In figure 3 are shown the 2 component and the taken as one part of the surface of zero potential. resultant fields in a shunt generator when the shunt In Figure 4 an attempt is made to visualize this cap coils occupy 50 per cent of the interpolar space. surface. In this figure 0 is the heart of the field, The horizontal field is the main and the vertical field the line DAQ is the neutral axis, and the line OADE is the compensating component. The resultant field the trace of the cap surface. It is the line required is circular in the shunt coils, and equilateral hyper- in that it is a level line inside the coil, an equipotenbolas in the insulation. The area of the triangle of tial outside the coil, and touches the neutral axis",
            " flux density on the yoke surface indicates the \"yoke\" On its inside surface and outside the coil the potential leakage flux, a component which apparently has is the full amount of the field coils. On its outside hitherto been neglected. Yet it should not be surface the potential is zero. As it is believed this neglected, as the saturation of the magnetic circuit is the first attempt to depict this surface for a dywill cause a small error in leakage flux to make a namo, the isometric projection is included in the larger error in ampere turns. Since the flux is dis- figure showing the cap surface. tributed along the yoke we may imaglne it as con- In figure 4 the effect of saturation in the core and centrated at a point half way between the pole edge yoke is shown by the core of the field being moved and the neutral axis.7 Moreover, the vertical component of the field determines the shape of the lines near the pole tip. Thus, referring to figure 3, note the dimensions, a, b, W, H, AH, the ampere turns of INTERPOLE the coil M, the maximum yoke density Bin, the yoke leakage flux 'k, the pole tip leakage flux caused by the compensating component of the field 4p, and s N N APp, its permeance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001380_sami.2015.7061901-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001380_sami.2015.7061901-Figure6-1.png",
        "caption": "Figure 6 Distribution of test plates on the platform (Isometric view)",
        "texts": [
            " Method of arranging the test plates on the platform The proposed method of arranging the test plates, with respect to the beginning of the coordinate system of the platform, derives from the assumption of changes of individual parameters (e.g. roughness), according to their mutual placement, with sufficient distance between the individual testing plates, or sectors, without a so-called \u201cinfluence zone\u201d. The arranging of the individual test plates into sectors labelled, i.e. TP01_SEC.01_P01 up through P05, TP01_SEC.02_..., TP01_SEC.03_..., TP01_SEC.04_..., TP01_SEC.05_..., TP01_SEC.06_..., TP01_SEC.07_..., TP01_SEC.08_..., TP01_SEC.09_..., is displayed in the following figure (Fig. 6, 7). E. Type of support material for a test plate Only one type of support material was used for all of the testing plates (i.e. from \u201cTP01_SEC.01_P01\u201d up through \u201cTP01_SEC.09_P05\u201d). Fig. 8 shows a sectioned view of a testing plate (specifically \u201cTP01_SEC.01_P01\u201d) together with a detailed view of the type of support material, or the characteristic shape of the so-called \u201cupper hatching teeth\u201d of the model/product. F. Manufactured test plates III. MEASURING OF SURFACE ROUGHNESS OF THE TEST PLATES IN THE INDIVIDUAL SECTORS This part of the work contains a description of the measurement of the surface roughness of the individual testing plates with respect to the sectors in which the testing plates were located"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003422_ecce.2016.7855360-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003422_ecce.2016.7855360-Figure1-1.png",
        "caption": "Fig. 1. Topology of magnetic-geared PM machine.",
        "texts": [
            " This paper mainly focuses on the power transferring feature of the magnetic-geared PM machines among the stator, modulation layer and rotor by proposed analytical equations. The equations of the back electromotive force (back EMF), electromagnetic torque and power are derived. In addition, the influence of the geometric parameters on the back EMF and electromagnetic torque are discussed. Finally, the analytical results are verified by the finite-element analysis (FEA) method. II. TOPOLOGY AND OPERATION PRINCIPLES Fig. 1 shows the topology of a inner rotor magnetic-geared PM machine. Different from the conventional PM machines, the magnetic-geared PM machine has a semi-closed stator and the dual rotation configuration of the rotor and modulation layer which is consisted of the steel blocks and nonmagnetic segments as shown in Fig. 1. The modulation layer, as a rotation part, is sandwiched between the stator and rotor of the conventional PM machine. In this machine type, there is the relationship of the PM pole pair numbers, modulation blocks numbers and winding pole pairs numbers as follows. r m sp Z p= \u00b1 (1) where ps and pr are the winding pole pair numbers and rotor PM pole pair numbers, respectively, and Zm is the numbers of the steel modulation block. To analyze the power transferring relationship among the rotor, modulation layer and stator of three phase magneticgeared PM machines, the open-circuit air gap flux density, winding function, back EMF and electromagnetic torque are firstly discussed in this sub-section"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000973_holm.2014.7031019-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000973_holm.2014.7031019-Figure2-1.png",
        "caption": "Fig. 2: Sketch view of a slip ring system",
        "texts": [
            " Wind generators are very complex systems and are composed of electrical and mechanical components. The installation of wind turbines in open area, will not only be exposed to weather-related hazards such as storms, lightning or freezing, but also to strong material stresses which cause serious damages. The most affected components are in the electrical generators. Our study is based on the double fed induction generators (DFIGs) those feed ac currents into both the stator and the rotor (Fig. 1). To connect the rotor to the network, slip ring apparatus will be used (Fig. 2). This allows power transmission from the rotor passing through a frequency converter into the electrical grid. By the transmission of the electrical currents through the resulting contact area between the two components (brush and slip ring), micro brush fires occur due to the sporadic contact separation. In this paper, we will analyze the background of these problems using diagnostic methods and develop an approach to mitigate them. II. PHYSICAL DESCRIPTION OF SLIP RING SYSTEM The transfer of the rotor current depends on the wear behavior in the slip ring system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002772_978-3-319-06590-8_181-Figure6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002772_978-3-319-06590-8_181-Figure6-1.png",
        "caption": "Fig. 6 Potential energy distribution diagram, Kb = 2e7 N/m; Kf = 3e5 N/m",
        "texts": [
            " However for moderate stiffness ratio foundation suspension stiffness increase brings to quick increase of critical speeds for almost all modes. In order to study dynamic of foundation structure on the base of abovementioned simple model, rotor bearings stiffness was set as 2E7 N/m, while stiffness ratio Kf/ Kb = 1.5E-02 was used for support structure. Mass ratio MF/MR = 25 was used to establish foundation mass. Undamped critical speeds and mode shapes are summarized in Table 1. Comparison of the Table 1 results and potential energy diagram (Fig. 6) had shown that first two modes are foundation rigid cylindrical and pivotal modes, due to concentration of energy in region of the supports, while the other modes were identified as pure rotor bending modes with more than 90 % of potential energy on the shaft. More advanced model was build for the same type of rotor but with plate type foundation structure (table size: 850 mm \u00d7 150 mm \u00d7 30 mm) using solid elements in ANSYS software, Fig. 7. Foundation base was made of steel with approximate weight \u224830 kg"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure3.5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure3.5-1.png",
        "caption": "Fig. 3.5 (a) A screw dislocation in a crystal. (b) Screw dislocation facilitates crystal growth by providing more bonding sites",
        "texts": [
            " The atoms above the dislocation line (highlighted by the \u201creversed T\u201d mark) are pushed together, whereas those below it are pulled apart, leading to regions of compression and tension above and below the dislocation line, respectively, as depicted by the shaded region around the dislocation line in Fig. 3.4b. Therefore, around a dislocation line, we have a strain field due to the stretching or compressing of the bonds. Another type of dislocation is the screw dislocation, which is essentially a shearing of one portion of the crystal with respect to another by one atomic distance (Fig. 3.5a). The displacement occurs on either side of the screw dislocation line. The circular arrow around the line symbolizes the screw dislocation. As we move away from the dislocation line, atoms in the upper portion become more out of registry with those below; at the edge of the crystal, this displacement is one atomic distance. The phenomenon of plastic or permanent deformation of a metal depends totally on the presence and motions of dislocations. Both edge and screw dislocations are generally created by stresses resulting from thermal and mechanical processing. A line defect is not necessarily either a pure edge or a pure screw dislocation; it can be a mixture (Fig. 3.5b). Screw dislocations frequently occur during crystal growth, which involves atomic stacking on the surface of a crystal. Such dislocations aid crystallization by providing an additional \u201cedge\u201d to which the incoming atoms can attach. To explain, if an atom arrives at the surface of a perfect crystal, it can only attach to one atom in the plane below. However, if there is a screw dislocation, the incoming atom can attach to an edge and thereby form more bonds, which is more favorable. In fact, many naturally formed crystalline materials are polycrystalline, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003961_physrev.47.781-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003961_physrev.47.781-Figure1-1.png",
        "caption": "FIG. 1. Series connection. FIG. 2. Parallel connection.",
        "texts": [
            " In attempting to compare more closely the vibrational patterns of the three plates by placing the three plates together in the simple interferometer and driving them through common electrode arrangements, the following noteworthy phenomenon was discovered: When two or more piezoelectric plates, possessing like resonance frequen cies and vibrational patterns, are driven through common electrode arrangements, the resonant responses are no longer those which are characteristic of each plate. Instead, they are characteristic of the plates as a group and are formed by the THE EDITOR twentieth of the preceding month; for the second issue, the fifth of the month. The Board of Editors does not hold itself responsible for the opinions expressed by the correspondents. \"electrocoupling\" of the mechanical modes of one plate with the mechanical modes of the other plates. In the series connection of Fig. 1 the three plates resonated in the cases of/mi and/1331 in unison at 386.5 kc/sec. and 556.3 kc/sec, respectively. The combination behaved as one plate. In / m i the plates tended to resonate separately, each at two nearby frequencies. The vibrational patterns associated with these two frequencies were slightly different from the pattern associated with the single frequency/1311 listed in Table I. In the parallel connection of Fig. 2 the plates tended, in the case of /1.331, to resonate separately, each at two fre quencies In the neighborhood of 556"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002252_1754337115577029-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002252_1754337115577029-Figure3-1.png",
        "caption": "Figure 3. (a) Schematic of the static measurement structure for evaluating deviation in ap-direction. (b) Measurement set-up of the JA:Ped3 system and PowerForce on the SRM ergometer.",
        "texts": [
            " The system is calibrated (linearity above 98% in all conditions \u2013 Figure 2) resulting in constant calibration factors for each pedal and each direction and does not require an additional calibration prior to each measurement. The shoe-side part of the click system was mounted on a beam to transmit the force directly on the pedal in a at Purdue University Libraries on July 9, 2015pip.sagepub.comDownloaded from cycling-specific way. The force was applied in the corresponding direction by different weights fixed to the beam (Figure 3(a)). The total load consisted of the beam, click system and weights. Weights of approximately 100N were applied in each direction, reaching a maximum weight of approximately 200N in the apand ml-directions and approximately 400N in vertdirection, which resembles the forces applied in both recreation and competitive sport.3,4,20 The mean of three measurements in each condition was compared to the applied load in absolute and percentage values. A negative absolute deviation implied that JA:Ped3 underestimated the pre-set load, while a positive absolute deviation implied overestimation of the system",
            "0 6 3.3 kg) familiar with click pedals were recruited. Tests were performed on an indoor ergometer (SRM; Schoberer Rad Messtechnik SRM GmbH, Germany), and to compare the output of the JA:Ped3, the ergometer was instrumented with (a) the SRM Powermeter (Schoberer Rad Messtechnik SRM GmbH), which provides information about the crank-torque without distinguishing between left and right force application and (b) the PowerForce system (O-Tec, Germany), which measures radial and tangential crank forces (Figure 3(b)). This configuration allowed a simultaneous data collection for all three systems. Both systems are reported to have acceptable accuracy.1\u20133 The mean error of the SRM Powermeter was reported to be 2.3% 6 4.9%1 and the mean error of the PowerForce system was shown to be 20.9% 6 4.1% for the tangential direction and 21.9% 6 6.6% for the radial direction.3 Prior to data capture, the PowerForce system was calibrated according to the instructions.3 Sampling frequency was 1000Hz for the JA:Ped3 and the PowerForce and 200Hz for the SRM Powermeter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003932_pi-4.1952.0037-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003932_pi-4.1952.0037-Figure2-1.png",
        "caption": "Fig. 2.\u2014Variation of energy with angle of separation; development of graphical method.",
        "texts": [
            "3 of f. 1 dt2 u)\\ between the limits 90 and 9X. M rxEx l12 \u2022 \u2022 \u2022 (ID (2) THE GRAPHICAL METHOD The fact that areas under torque/angle curves are used in the foregoing solution suggests a method in which the problem is treated on an energy basis in the following way- Evaluating the integrals in eqn. (10) from 90 to 0 we obtain \u2022 (12) We now plot (rxExE2/coXx2)(cos 90 \u2014cos9) and {r2ExE2/a)Xx2) (cos 90 \u2014 cos 9) against 9 for 9Q \u2014 0 and 9 going from 0 to TT. These curves are shown as (a) and (b) in Fig. 2. Select the point (a) - ' - ' 2 (cos Oo - cos 0). <oAi2 (cos Go - cos G). P corresponding to 90 on curve (a) and draw the straight line PQ of slope PM/co; the intercept, such as ya at 9a, between the line and the curve then gives the kinetic energy - ( \u2014,\") 2 \\ at J We have now represented graphically part of eqn. (10), namely sin 9 dO where yx is the intercept when 9 \u2014 9V The clearing of the fault cannot produce an instantaneous change in the kinetic energy; thus beyond 9t Trd6\\2 sin 9d9 x2 which may be represented by adding RS = ^j to the ordinate of curve (b) at 9X and drawing from S the straight line ST of slope PMla> (i",
            " The curves for the cosine expressions can be plotted for 0O = 0 and 0 going from 0 to 77 and the straight lines of slope PM/(O drawn as before, from the 0O point on the lower curve and tangential to the upper curve. The procedure for determining the critical switching time is now the same as in the case for the symmetrical machines without resistance. It may be well to remark that for transient conditions B and B' are generally negative; thus the curves for the cosine expressions are lower 1 - {xd - xq)( 0 xq + 2x sin2 - xdxq + 2x(xq sin2 - + xd cos2 towards the origin and higher in the region where 0 = rr/2 than the corresponding curves of Fig. 2. (5) ACKNOWLEDGMENTS The author wishes to express his thanks to Professor A. E. Green for advice and help in the mathematical statement and to Dr. Mid Ouyang for working out the values for Fig. 4. He is also indebted to Sir Isaac Pitman and Sons, Ltd., for permission to reproduce data from H. Rissik's book \"Power System Interconnection,\" for which they hold the copyright ALTERNATORS: A GRAPHICAL SOLUTION OF THE TWO-MACHINE CASE 371 (6) REFERENCES (1) PARKER, R. H., and BANCKER, E. H.: \"System Stability as a Design Problem,\" Transactions of the American I"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000075_gt2012-68354-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000075_gt2012-68354-Figure4-1.png",
        "caption": "Figure 4. The Kevlar brush seal showing the pyrometer port where the quartz glass rod was adapted.",
        "texts": [
            " The metallic brush seal has had bristles with a radial length of 6.5mm and the Kevlar seal has had fibers with a radial length of 5.5mm. Their axial lengths were 2mm and 3mm respectively. The major instrumentation setup can be summarized as follows: 1) 2 pyrometers for transient recording of the rotor material temperature at the contact zone (bristles/rotor, location A) and at the rotor tip (location B). The pyrometer which was viewing location A was immersed into the bristle package of the brush seal (Fig. 4) at a distance of 1mm from the rotating part. Its tip was elongated using a quartz glass rod. The other pyrometer was viewing location B through a glass window in the bearing chamber flange (Fig.3). The 2 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 05/05/2014 Terms of Use: http://asme.org/terms measuring and calibration technique will be explained in detail in the next chapter. 2) pressure, temperature and mass flow measurements of the sealing air upstream of each seal individually"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003857_0954406215589843-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003857_0954406215589843-Figure4-1.png",
        "caption": "Figure 4. Rigid body rotor model.",
        "texts": [
            "001 rad), and therefore the rotation matrix from e to ep can be written as epx epy epz 2 4 3 5 \u00bc 1 0 0 1 1 2 4 3 5 ex ey ez 2 4 3 5 \u00f02\u00de With this relation between the frame orientations, and the radial distance r between the origins of frames e and ep (see Figure 2), the surface point coordinates in the bearing centre inertial frame become x y z 2 64 3 75 \u00bc 1 0 0 1 1 2 64 3 75 x p y p z p 2 64 3 75 0 0 r 2 64 3 75 \u00f03\u00de And the velocity components are given by x _ y _ z _ 2 64 3 75 \u00bc 0 0 _ 0 0 _ _ _ 0 2 64 3 75 x p y p z p 2 64 3 75 0 0 _r 2 64 3 75 \u00f04\u00de An example model of the motion of the rigid rotor shaft is shown in Figure 4. If motion along the axial direction ey is restricted and the rotor has a circular cross section, the position of the rotor surface is fully determined by the four coordinates shown in the figure . Translation X along ex . Translation Z along ez . Rotation around ecx . Rotation around ecy The translations and rotations are assumed to be very small, so the order in which rotations are applied does not matter. The rotor shaft has a circular cross section; therefore, rotation around its axis does not change topology"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001478_s12206-014-0314-0-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001478_s12206-014-0314-0-Figure2-1.png",
        "caption": "Fig. 2. Loading device for X-ring under uniform squeeze rate and internal pressure.",
        "texts": [
            " The molding box was heated and cooled according to the molding cycle [18]. The X-ring made from high temperature epoxy resin was then taken out of the molding box, and stress freezing was carried out. The X-ring fabricated from high temperature epoxy resin, the cylinder, and the guide-ring was heated for about 1 hr. at a temperature of 120\u00b0C in the stress freezing furnace. The Xring, cylinder, and guide ring were then assembled in the furnace to obtain the required squeeze and internal pressure. The assembled loading device (Fig. 2) was then heated according to the stress freezing cycle [19]. The stress frozen Xring was removed from the stress freezing furnace and sliced. 2 mm slices were cut at 120\u00b0 intervals from the stress frozen X-ring. The slices were polished until their thickness was approximately 0.8 mm. The finished slices were placed in a glass box with a mixed solution of \u03b1-bromonaphthalene and fluid paraffin at a volume ratio of 1:0.585. The glass box with the finished slices was placed on the transparent photoelastic experimental device"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003394_icinfa.2016.7831871-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003394_icinfa.2016.7831871-Figure1-1.png",
        "caption": "Fig. 1. An industrial XY linear-motor-driven motion stage.",
        "texts": [
            " Remark 3: It is shown in [8] that A is positive semi-definite, if and only if the following condition is satisfied: ks1 \u2265 k2 + \u03b81k1 \u2212 \u03b82 \u2212 \u03b83gf \u2212 \u03b8T4 gh2 + 1 2\u03b81k31 [\u03b82k1 + k1(\u03b83gf + \u03b8T4 gh2) + |\u03b8T4 gh1|]2 (24) Theorem 2: If the DCNNARC law in (18) is applied, then we can get the following conclusions [8]: 1) In general, all signals are bounded. Furthermore, the positive definite function V \u2032 s = 1 2Mp2 + 1 2Mk21e 2 is bounded by V \u2032 s \u2264 exp(\u2212\u03bb\u2032t)V \u2032 s (0) + \u03b5\u2032 \u03bb\u2032 , \u03bb \u2032 = min{ 2k2 \u03b81max , k1} (25) 2) If after a finite time t0, there exist parametric uncertain- ties only (i.e., d\u0303 = 0, \u2200t \u2265 t0), then, in addition to above result 1), zero final tracking error is also achieved, i.e., e \u2192 0 and p \u2192 0 as t \u2192 \u221e. The proposed control scheme is tested on the X-axis of an industrial XY iron-core linear-motor-driven motion stage showed in Fig. 1. The proposed control algorithms are implemented in Matlab/Simulink environment and the real-time code is automatically operated in a dSPACE DS1202 control system. The position sensor is a linear encoder with resolution of 156.25 nm after quadrature. And the controller board executes algorithms at a sampling frequency of fs = 5 kHz. System identification has been conducted in our previous work [1], and standard least-square identification is performed to identify the physical parameters of the linear motor stage"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000968_20140824-6-za-1003.01643-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000968_20140824-6-za-1003.01643-Figure3-1.png",
        "caption": "Fig. 3. A portion of UAV flight corridor",
        "texts": [
            " Here, a number of definitions must be made: Definition 3: Consider the path segment corresponding to the Cruise state of the UAV. Flight corridor is the union of all (right and left) loitering circles tangent to the cruise segment of the flight. It is clear that if the flight corridors of two UAVs do not intersect, there is no possible collision event between those two UAVs. Definition 4: The smallest circle that encompasses the right and left loitering circles is called the safety circle. In Fig. 3, part of the flight corridor of a UAV and its safety circle is schematically shown. It is clear that the UAV can stay in its safety circle, without violating the constraints in (2) and (3). In other words, safety circle of the UAV can stay stationary, if required. Definition 5: A collision avoidance maneuver is when a UAV moves into its right loitering circle and finishes one complete circle. Once the UAV completes one full loitering circle, it is back on its original cruise path toward the objective circle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001597_roman.2015.7333638-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001597_roman.2015.7333638-Figure1-1.png",
        "caption": "Fig. 1. View from above of a swarm of 6 quadrotors in a wedge formation. Effect of the concurrent behaviors following user commands (forward) and keeping the formation shape on the robot i.",
        "texts": [
            " The ideal position can be written in world coordinate as: xi = xom + ri cos(\u03b1i \u2212 \u03b1h) yi = yom + ri sin(\u03b1i \u2212 \u03b1h) (1) zi = zom where xom, yom and zom are the positions of the center of mass of the formation, ri is the module of the distance between the robot i in its ideal position and the center of mass, \u03b1i is the angle between the straight line corresponding to the heading of the swarm and the segment connecting the center of mass with the robot i and and \u03b1h is the angle between the straight line corresponding to the heading of the swarm and the x axis. The effect of user commands on the robot i can be represented, in world coordinates and in relation to the user input (forward or backward), as a vector vui of three components in space: vuix = \u00b1Gu cos(\u03b1h) vuiy = \u2213Gu sin(\u03b1h) (2) vuiz = 0 where Gu is a gain factor related to the space displacement to be travelled each step and its sign is related to the user command (forward or backward). Fig. 1 gives a schematization of the effect of these two concurrent behaviors. When the human operator imposes a rotation of the formation, the vector vui is calculated as: vuix = G\u03b1(ri cos(\u03b1i \u2212 \u03b1h \u00b1 \u03b1k)\u2212 ri cos(\u03b1i \u2212 \u03b1h)) vuiy = G\u03b1(ri sin(\u03b1i \u2212 \u03b1h \u00b1 \u03b1k)\u2212 ri sin(\u03b1i \u2212 \u03b1h)) (3) vuiz = 0 being \u03b1k the angle displacement controlled by the user and G\u03b1 a further gain factor related to the rotation. The orientation of the single agents follows the heading of the swarm. If the position of a robot falls inside the sphere of influence of one of the obstacles, a further vector vob should be added, directed from the center of the obstacle to the center of mass of the formation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000589_pc.2015.7169967-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000589_pc.2015.7169967-Figure2-1.png",
        "caption": "Fig. 2. Illustration of the Pendubot\u2019s equilibrium positions with regards to the arrangement of its links. From left to right: down-down \u21ca, down-up , up-down , up-up \u21c8.",
        "texts": [
            " Taking into account viscous friction in both joints the dynamic system (2) can be rewritten as follows: q\u03081= 1 \u03b81\u03b82\u2212\u03b823cos 2q2 [\u03b82\u03b83sinq2(q\u03071+q\u03072)2+\u03b823cosq2sin(q2)q\u030721 \u2212\u03b82\u03b84gcosq1+\u03b83\u03b85gcosq2cos(q1+q2)+\u03b82\u03c41\u2212\u03b82b1q\u03071 +(\u03b82+\u03b83cosq2)b2q\u03072] (3) q\u03082= 1 \u03b81\u03b82\u2212\u03b823cos 2q2 [\u2212\u03b83(\u03b82+\u03b83cosq2)sinq2(q\u03071+q\u03072)2 \u2212(\u03b81+\u03b83cosq2)\u03b83sin(q2)q\u030721+(\u03b82+\u03b83cosq2)(\u03b84gcosq1\u2212\u03c41) \u2212(\u03b81+\u03b83cosq2)\u03b85gcos(q1+q2)+(\u03b82+\u03b83cosq2)b1q\u03071 \u2212(\u03b81+\u03b82+2\u03b83cosq2)b2q\u03072]. (4) These are Pendubot\u2019s non-linear equations of motion which are to be used for the design of the nonlinear predictive controller. By introducing the state vector x(t) = [x1, x\u03071, x2, x\u03072]T =[q1, q\u03071, q2, q\u03072]T and denoting u = \u03c41, the equations of motion of the Pendubot may be re-written in the following form: \u23a7\u23aa\u23aa\u23a8\u23aa\u23aa\u23a9 x\u0307 = f(x,u, b) y = h(x, b) (5) Recall that for u = 0 the system (5) exhibits in total four main equilibrium positions, see Fig. 2, the first being a stable and the other three unstable ones. From these, challenging from the control standpoint are mainly the down-up (further ) and upup (\u21c8) Pendubot\u2019s unstable positions. In the scope of this work is a point-to-point transition task between different equilibria by exploiting the nonlinear dynamics (5) subject to constraints, as proposed in the following sections. The parameters such as mass mi, the pendulum arm length li, and the distance lci of the center of mass to the corresponding joint are directly measurable, and read as m1 = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure6.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure6.4-1.png",
        "caption": "Fig. 6.4 Die casting process",
        "texts": [
            " This process can be applied to a wide range of material sizes, ranging from small scales (~10 g) to very large scales (~1000 kg). The process induces homogeneous material properties regardless of directions or positions. Cast production is costly, but such production requires only one process to form a complex body shape. Hence, the overall process is still economical for mass production. Furthermore, as the excessive material around the casted body can be trimmed, remelted, and reused in the following casting processes, there is no material wastage, ideally. Die casting (Fig. 6.4) is one type of casting process involving the usage of molds called die. In die casting, metal is usually first melted. Then, it is injected into the die cavity by a plunger under high pressure (~1000 to 20,000 psi). The cavity is formed by two hardened die with the desired geometry. Once the die cavity is filled, the molten metal is allowed to cool down and solidify inside the die. After a solid product is formed, it is removed from the die by being pushed out by ejection pins. Sometimes, there may be trimming following the removal to remove any sprue surrounding the final product to finish the whole process"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002311_imece2016-65745-Figure14-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002311_imece2016-65745-Figure14-1.png",
        "caption": "FIGURE 14: CONTACT STATUS IN P3G PROFILE",
        "texts": [
            " The interference fit has the lowest stress and is the recommended fit for high torsional bending load. The reason for the lower stress in interference fit is relatively larger contact area compared to other fits. The expansion of the hub as a result of the stress also causes the interference to decrease and subsequently decrease the contact stress. 5 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90996/ on 07/24/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Fig. 14 shows the contact status of the P3G polygonal shaft and hub connection. It can be observed that the contact area is small in clearance fit and looks like a rectangular strip. Due to the expansion of hub by the applied stress, the contact region in transition fit is also localized, although it is larger than clearance fit. Similar results can be observed from the contact status for P4C polygonal shafts from Fig. 15 for clearance and transition fit. The contact status for clearance fit is similar to the results found by Czyzewski and Odman [5] for pure torsional plane stress condition"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000409_978-3-642-28572-1_38-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000409_978-3-642-28572-1_38-Figure3-1.png",
        "caption": "Fig. 3 Exploration of an environment including an irregular shape. (0) Surface to be explored, composed of geometric primitives and a generic, rough surface. (1) Beginning of the exploration: the rough surface is represented as a mesh. (2) A primitive is discovered. (3) End of the exploration. The primitives require much fewer points than the mesh.",
        "texts": [
            " The combination of the strategies described above resulted at least twice as fast as a random or partially random approach, as well as any of the two steps of the strategy used separately. Simulations with rough and irregular surfaces demonstrate the robustness of the mapping algorithm. The primitives are recognized in the presence of high surface noise, even though they require more touch points. Primitive deformations are treated as intuitively expected: low deformations are tolerated, while higher deformations force a primitive to be represented as a composition of similar primitives or as a mesh. Figure 3 shows the exploration of an environment composed of both simple primitives and an irregular surface. Simulations with surrounding viscous fluids showed that, for the speed used in our research, fluid effects do not significantly affect the approach [23, 24]. The main effect of high temperature on the manipulator is the induced thermal expansion of the gears. A considerable joint backlash is required in a down-well manipulator so that the large swings in temperatures do not cause the joints to bind up"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.32-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003618_b978-0-08-099425-3.00006-6-Figure6.32-1.png",
        "caption": "FIGURE 6.32",
        "texts": [
            " As discussed in Chapter 3 single-purpose general-purpose programs such as ADAMS/Car include a driveline model as part of the full vehicle as a means to impart Loss in velocity as vehicle \u2018coasts\u2019 through the lane change manoeuvre. Simple drive torque model. REV, revolute joint. torques to the road wheels and hence generate tractive driving forces at the tyres. Space does not permit a detailed consideration of driveline modelling here but as a start a simple method of imparting torque to the driven wheels is shown in Figure 6.32. The rotation of the front wheels is coupled to the rotation of the dummy transmission part as shown in Figure 6.32. The coupler introduces the following constraint equation: s1.r1 + s2.r2 + s3.r3 = 0 \u00f06:17\u00de where s1, s2 and s3 are the scale factors for the three revolute joints and r1, r2 and r3 are the rotations. In this example suffix 1 is for the driven joint and suffixes 2 and 3 are for the front wheel joints. The scale factors used are s1\u00bc 1, s2\u00bc 0.5 and s3\u00bc 0.5 on the basis that 50% of the torque from the driven joint is distributed to each of the wheel joints. This gives a constraint equation linking the rotation of the three joints: r1 = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002889_amm.851.273-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002889_amm.851.273-Figure1-1.png",
        "caption": "Fig. 1 The bending-torsion-axis\u2013pendulum dynamic analysis model of the helical gear.",
        "texts": [
            " By considering the meshing characteristics of helical gear, and then combined with the premise of the influence factors of helical gear transmission system, the bending-torsion-axis-pendulum coupling nonlinear analysis model of helical gear is established. The corresponding dynamic differential equations are derived and the dynamic differential equations are solved by numerical method, and to find the basic rules of dynamic characteristics of helical gear drive system. The bending-torsion-axis\u2013pendulum dynamic analysis model of the involutes helical gear transmission system which is established by the lumped mass method is shown in Fig. 1. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (#74468567, Iowa State University, Ames, USA-20/02/17,04:12:38) Each gear has five degrees of freedom of space from the model: along the X and Y direction of the gear shaft; Z axial direction; \u03b8z direction; \u03b8y direction around the Y axis. It has known helical gear meshing forces and time-varying meshing stiffness kh, meshing damping cm, meshing error e"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003755_978-3-319-28451-4-Figure13.4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003755_978-3-319-28451-4-Figure13.4-1.png",
        "caption": "Fig. 13.4 Power-load element and an ideal voltage source",
        "texts": [
            "2 Power-load element P and its hyperbolic characteristic 362 13 Power-Source and Power-Load Elements Note The offered graphic representation of both power-load \u00feP and power-source P elements corresponds by analogy to a voltage source and source of current. Further, we will consider power-load elements using the results [5, 6]. 13.3 Influence of Voltage Source Parameters and Power-Load Element onto a Power Supply Regime 13.3.1 Ideal Voltage Source The power-load element P with an ideal voltage sourceV0 is represented in Fig. 13.4. A volt-ampere characteristic of P is the known hyperbola I \u00bc P=V . In turn, a volt-ampere characteristic of V0 represents a vertical line with the coordinate V \u00bc V0. A point of intersection \u00f0I;V0\u00de of this line and hyperbola determines a single or one-valued regime of this circuit. For a different voltage source value V1 0 , a new point \u00f0I1;V1 0 \u00de will be the energy equivalent point to the initial point \u00f0I;V0\u00de because this circuit does not possess the own or internal scales, which can specify the qualitative characteristics of an operating regime"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003178_icmid.2016.7738928-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003178_icmid.2016.7738928-Figure4-1.png",
        "caption": "Fig. 4. Base plate Structure",
        "texts": [
            " The roller is 20 mm in diameter and rotated with a dedicated DC servomotor. As the whole motorroller system is placed on linear stage driven by a stepper motor, rotation and traversing speed and/or direction of the roller can be controlled independently. In this research, typical counter roller mode was used. Joule heaters installed in the build chamber can raise the powder bed temperature up to 200 \u02daC. The same material as the powder is injection molded, machined and shaped to prepare a base plate as illustrated in Fig. 4. To reinforce the base plate, an aluminum alloy backup plate is fixed with screws. Previously, it was reported that violent reactions such as sparkling and fuming were caused during low temperature process of high temperature plastic [3]. In the report, these reactions can be suppressed when intensity of the laser dedicated to sintering is reduced by expanding beam spot and lowering laser power. In present research, two beam spot diameters 130 \u03bcm and 730 \u03bcm were used. The former one was the minimum value of the apparatus employed in this research"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000473_978-3-319-20463-5_8-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000473_978-3-319-20463-5_8-Figure5-1.png",
        "caption": "Fig. 5 Run-up pre-processed vibration signal (left) and test rig scheme [8] (right)",
        "texts": [
            " The measuring dynamic range was from 0 to 1800 rpm during 4 s, starting in steady state and reaching maximum operating condition, with a sampling frequency of 20 kHz. Vibration signal is preprocessed by means of a 9th order low-pass Chebyshev filter with cut frequency at 1.2 kHz, and down-sampled to 2.5 kHz. Considering that run-up test goes until 1800 rpm or 30 Hz. Inasmuch as the maximum speed is similar to the simulated signal case, the work interval used is [35, 300]Hz, and the signal is once again low-pass filtered to 300 Hz and down-sampled to 625 Hz. The STFT has 1024 frequency bins and a Hanning window of 256 points, and the parameter d \u00bc 1Hz: The Fig. 5 shows the pre-processed vibration signal (left) and the test rig scheme (right). In this experiment the reference shaft speed was not measured; therefore, the IF extraction performance is defined using the two proposed angle-order map stationarity measures. A visual validation is performed in the obtained angle-order map by extracted IF. Figure 6 shows the minimum qi for each partition Pi and all variability values obtained at 5th and 7th partition using Kurtosis-CV (left) and PCA index (right) respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002154_1077546314557851-Figure11-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002154_1077546314557851-Figure11-1.png",
        "caption": "Figure 11. (a) 2DOF model representing the rotor bearing set up and (b) generation of excitation force vector for radial vibration measurement.",
        "texts": [
            " non-linearity due to the Hertzian force/deformation relationship, the varying stiffness resulting from load transmission via a finite number of rolling elements, the presence of clearance between the rolling elements and the bearing races, and the effect of lubricant film. The two parts generate simultaneous deterministic as well as stochastic response vectors fxdg and fxsg which superpose to make the total response fxg. Figure 10 shows a sketch of the test rig along with the instrumentation used and Figure 11(a) is the equivalent 2DOF model for the same. The first natural frequency of the rotor bearing system as found from the at TEXAS SOUTHERN UNIVERSITY on November 19, 2014jvc.sagepub.comDownloaded from impact test is approximately 492Hz, and the theoretical calculation of natural frequency based on 2DOF system yields a value of 513.5Hz which implies that 2DOF model is able to predict the actual value within an error less than 4.36%. Hence, a simple 2DOF model was considered sufficient to represent the system dynamics. For the 2DOF model shown in Figure 11(a), equation (7) may be rewritten elaborately in equation (8) m1 0 0 m2 \u20acx1 \u20acx2 \u00fe c c c c _x1 _x2 \u00fe k1 \u00fe k2 k2 k2 k2 x1 x2 \u00bc f1\u00f0t\u00de f2\u00f0t\u00de d \u00fe f1\u00f0t\u00de f2\u00f0t\u00de s \u00f08\u00de aswhere, \u2018m1\u2019 is the mass of extended portion of the shaft and inner race, \u2018m2\u2019 is the mass of outer race and housing, \u2018c\u2019 is the damping coefficient of balls, due to oil film that builds up during rotation; magnitude of c lies between 0.25e-5klin to 2.5e5klin where klin is the linearized stiffness of the bearing in N/mm, \u2018k1\u2019is the stiffness of the shaft (on which bearing under investigation is mounted), \u2018k2\u2019 is the linearized bearing stiffness",
            " Following Choudhury and Tandon (2006) k2 \u00bc Ptotal= max 0:5Pd; Ptotal \u00bc Kd 3=2 max P 1 \u00bc 1 \u00bd1 \u00f01=2\"\u00de \u00f01 cos \u00de 3=2 cos ; where \" is the load distribution factor, \" \u00bc 1 2 \u00f01 Pd 2 r \u00de; Pd is the diametric clearance and r is the radial shift of the ring at \u00bc 0 and 1 is the extent of load zone; Kd \u00bc \u00bd 1 1=Ki\u00f0 \u00de 1=n \u00fe 1=Ko\u00f0 \u00de 1=n n; Ki=o \u00bc 2:15 105 X 1=2i=o \u00f0 i=o\u00de 1:5 \u00f0N=mm1:5\u00de: The vectors ffi t\u00f0 \u00deg d (i\u00bc 1, 2) denote the excitation force vector due to a localized defect acting on mass mi at time t. The generation of the excitation force causing radial vibration is explained in Figure 11(b) and is obtained as fi t\u00f0 \u00de d \u00bc f(t)P( )cos . The generated pulse function f(t), has been represented by a rectangular pulse form whereas P\u00f0 \u00de is the load at the point of excitation. 4.2.1. State equation. Considering the phase variables as states, the four state variables of the 2DOF model of rotor bearing system (Figure 11(a)) are given by X1 \u00bc x1, X2 \u00bc _x1, X3 \u00bc x2, X4 \u00bc _x2; _X1 \u00bc _x1 \u00bc X2, _X2 \u00bc \u20acx1, _X3 \u00bc _x2 \u00bc X4, _X4 \u00bc \u20acx2; where x1 and x2 have been explained in Figure 11(a). Therefore, the state vector fX\u00f0t\u00deg can be written as fX t\u00f0 \u00deg \u00bc fX1X2X3X4g T \u00bc fx1 _x1x2 _x2g T: Taking ts as the sampling time, equation (8) can be written in discrete state space form as shown in equation (9) where wi k\u00f0 \u00de i \u00bc 1, 2, 3, 4\u00f0 \u00de represents process noise corresponding to the ith state. The equation (9) follows the form of the state equation discussed in the discrete Kalman filter section. 4.2.2. Measurement equation. Lack of full observability could prevent the state estimators (Kalman and H1 filters here) from successfully estimating the states (Simon, 2006)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000218_978-3-319-52219-7-Figure3.17-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000218_978-3-319-52219-7-Figure3.17-1.png",
        "caption": "Fig. 3.17 (a) Top view and, (b) side view of the optical table for vertical alignment",
        "texts": [
            " In the next step, it is aligned such that the optical axis is in plane of the incident beams, in the direction pointed by the incident beams. In this section we will describe an alignment method based on Lee\u2019s two step method [58], but replacing the He-Ne laser, the plane mirrors and the iris with simple laser diodes and alignment scales. The alignment procedure is as described in Lee\u2019s work in [58] but adapted to use collimated laser diodes and fixed apertures instead of the He-Ne laser with beam spliters and plane mirrors. Figure 3.17 shows the basic arrangement of the laser diodes (LD1, LD2) and off-axis parabolic mirror (OAPM) to be aligned. A typical 635 nm collimated laser diode modules with a round beam size of 2\u20133 mm can be used for the laser sources. For alignment, the lasers can be mounted on a Kinematic Mount. The two-laser beams (Beam 1 and Beam 2) are made parallel to the surface of the optical table using a set of fixed apertures ( 2\u20133 mm), that are set at the same 62 3 Experimental Methods height. Alternatively, Magnetic Beam Height Measurement Tool with alignment holes can be used to make the lasers parallel to the optical table surface. The laser beams are also made parallel to each other. If using fixed apertures mounted on posts in the optical slots, the beams should automatically get parallel to themselves as well. The distance between the beams x is determined by the width of the OAPM. It can be shown that the optical axis of OAPM get aligned better with the incident beams for larger x [58]. In this section we use the same rectilinear coordinate system as used in Lee\u2019s method [58], as shown in Fig. 3.17. In the first alignment, the vertical alignment, the optical axis of the OAPM is placed in the plane created by the two incident beams. In the next alignment, the horizontal alignment, the optical axis is made parallel to the incident beams. The focal point of the OPAM is found where the two reflected beams meet. In order to bring the optical axis in plane of the incident beams, the steps for the vertical alignment procedure are: 1. The OAPM is placed on the optical table using kinematic mounts and the optical axis is made approximately parallel to the incident beams as shown in Fig. 3.17a. 3.3 Test and Characterization 63 2. Then the reflected beam 1 is made parallel to the optical table. The same fixed apertures can be used at two different locations to check the parallelism. A Magnetic Beam Height Measurement Tool with alignment holes can be used as well. 3. Next we make the reflected beam 2 parallel to the optical table using the same method as used above for the reflected beam 1. 4. If the reflected 2 beam moves upward, then we decrease the height of the OAPM. 5. Similarly, if the reflected beam 2 moves downward, we increase the height of the OAPM (see Fig. 3.17b). Steps 2\u20135 are repeated till the both the reflected beams are parallel to the surface of the optical table, thus bringing the optical to the plane of the incident beams. Next, in order to make the optical axis parallel to incident beams, the steps for the horizontal alignment procedure are: 1. A plane mirror, preferably mounted on a kinematic mount, is placed at the intersection of the two reflected beams. This is not necessarily the focal point since the optical axis may not be parallel to the incident beams yet"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003547_iciinfs.2016.8262981-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003547_iciinfs.2016.8262981-Figure1-1.png",
        "caption": "Fig. 1. Two dimensional aerofoil",
        "texts": [
            " Also, the a priori knowledge of the bound of the disturbances is not required. The remaining paper is organized follows. In section II the model of two dimensional aerofoil is presented. Controller design using UDE based SMC approach is presented in section III. Section IV presents the numerical simulations and comparative performance of SMC and UDE-SMC based controllers. Section V concludes the work. A schematic diagram of a prototypical 2-dimensional aero elastic aerofoil with a trailing-edge control surface is shown in Fig.1. The governing equations of motion are given as [12] and [13],[ \ud835\udc5a \ud835\udc46\ud835\udefc \ud835\udc46\ud835\udefc \ud835\udc3c\ud835\udefc ] [ \u210e\u0308 ?\u0308? ] + [ \ud835\udc3e\u210e 0 0 \ud835\udc3e\u210e ] [ \u210e \ud835\udefc ] = [ \ud835\udc3f(\ud835\udefc, \u210e\u0307,?\u0307?, \ud835\udefd) \ud835\udc40(\ud835\udefc, \u210e\u0307, ?\u0307?, \ud835\udefd) ] (1) Where \ud835\udc5a is the mass of aerofoil, \ud835\udc46\ud835\udefc is the static moment about the elastic axis, \ud835\udc3c\ud835\udefc is the inertia moment about the elastic axis, \u210e and \ud835\udefc are the plunging and pitching coordinates, respectively, \ud835\udc3e\u210e and \ud835\udc3e\ud835\udefc are the stiffness of \u210e and \ud835\udefc respectively, \ud835\udefd is the deflection of the flap. \ud835\udc4e is dimensionless distance from mid chord to elastic axis and \ud835\udc4f is the semi chord"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003119_icelmach.2016.7732831-Figure10-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003119_icelmach.2016.7732831-Figure10-1.png",
        "caption": "Figure 10: Rotor dynamic eccentricity when bearing cage fault occurs",
        "texts": [
            " The system response function obtained with the impulse function out of Figure 8 and the system parameters mrot,DE, krad,DE and crad,DE is presented in Figure 9. This way of obtaining the AMB set points is less time-consuming than FEM and will be used for the calculation of both inner race and outer race bearing fault rotor movements. However, FEM can still be used to determine the exact pulse duration and magnitude of the impulse functions for different severity stages of the pits. When a bearing cage fault occurs, the bearing balls are not equally divided around the inner ring (Figure 10). Sequentially, the rotor is subjected to a dynamic eccentricity at a frequency equal to the FTF (fFTF). Therefore, the equivalent rotor displacement y(t) and x(t) can be expressed as: y(t) = Acage \u00b7 sin (2 \u00b7\u03c0 \u00b7 fFTF \u00b7 t) (11) x(t) = Acage \u00b7 sin ( 2 \u00b7\u03c0 \u00b7 fFTF \u00b7 t+ \u03c0 2 ) (12) The magnitude Acage in the expression can be used as an indicator for the fault severity. Because the bearing cage fault set point vary relatively slow in time with respect to an outer race bearing fault (fFTF 7Hz for the DUT at a rotational speed of 1460rpm), the focus in dimensioning the control system of the AMB will lie on emulating a bearing outer race fault"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002311_imece2016-65745-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002311_imece2016-65745-Figure2-1.png",
        "caption": "FIGURE 2: FORCES IN A POLYGONAL CONNECTION",
        "texts": [
            " Cartesian Form of Equation:              )(sin)(nsin)-ne(cos)(ncos-e 2 d )= x( m              )(cos)(nsinne)(sin)(ncos-e 2 d )= y( m Polar Form of Equation: 2 2 m ))nsin(ne()ncos(e 2 d )(r                    )ncos(ed5.0 )nsin(ne tan)( m 1 The P3G is a harmonic curve as described by Eqn. (1) or (2), while the P4C is the superimposition of the four lobe profile as described by Eqn. (1) or (2) and grinding diameter circle as shown in Fig. 1 [8, 9], where d1 is the grinding or outer diameter and d2 is the inner diameter. Referring to Fig. 2, the outer profile is the hub and the inner profile is the shaft. For an idealized point of contact at a position P with a polar coordinate (r, \u03d5), the force acting at the point P in shaft hub connection can be expressed as two orthogonal components FTP (tangent force on profile) and FNP (normal force on profile) with their lever arms t and N. Considering the drive from a gear, the resultant force of FTP and FNP is equal to the tangential force, FTG from gear. The radial force from the gear is FRG"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000670_kem.651-653.901-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000670_kem.651-653.901-Figure1-1.png",
        "caption": "Fig. 1 Example of a joint cross section obtained via profiles extracted through Leica DCM 3D: a) Profile extracted from the weld surface; b) Profile extracted from the lower surface.",
        "texts": [
            "75 kW-25 mm/s; 2 kW-25 mm/s), lack of penetration occurred due to an insufficient power density for the specimens realized at 1.5 kW and due to an insufficient heat input in the others. Successful combinations of welding speed and laser power are reported in Table 1, hence a reduced factorial plane 2 2 was carefully studied. After welding, for each specimen both the weld face and the root surface were acquired using a Leica DCM 3D confocal microscope. For each surface acquired with the software LEICA MAP 7.0, five profiles were extracted and the measures, showed in Fig.1, were carried out. After these acquisitions, microstructure and microhardness of the joints with different welding parameters were investigated. Then, from each weld, three metallurgical specimens, with the main surface perpendicular to the welding direction, were cut away from the welded sheets to record microscopic weld geometry, microstructure and to perform the microhardness test. Every specimen was mounted in a proper thermoset resin, polished with grinding discs and etched by HF solution.The optical observations and microhardess measures were carried out using a SEM digital camera respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002683_978-3-319-11930-4-Figure4.16-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002683_978-3-319-11930-4-Figure4.16-1.png",
        "caption": "Fig. 4.16 Autopsy data for RCF elements in configuration 1 of Table 4.1 operated at 3.61 GPa loading. Suspension after 1.7 h of rotation at 130 Hz, accounting for approximately 7.8 105 rod stress cycles",
        "texts": [
            " If allowed to continue without lubrication, the increased friction would accelerate the onset of subsurface spall of the ball and rod. Figures 4.16, 4.17, and 4.18 reveal interaction between the balls, the rod, and races, at different stages during testing. Post-test autopsy of the rotating parts is extremely valuable and may be used to guide future testing. Mechanisms of wear and material transfer may be investigated for the purpose of validating a model or analysis technique to the test data. For example, the test in Fig. 4.16 was autopsied after just 1.7 h, and results confirm that the elements were in the full-film lubrication regime. There is sufficient silver film present such that asperity-to-asperity contact is less significant. The results in Fig. 4.17 show evidence of mixedboundary conditions on the contact surfaces. Some of the silver film has been pushed out of the wear track, rendering it useless to the rotating parts. In fact, the film in Fig. 4.17 is very near total depletion, indicating that if allowed to continue the contact would begin entry in the boundary layer regime and rapid failure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure5.1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure5.1-1.png",
        "caption": "Fig. 5.1 Schematic of four regions on the fracture surfaces of glassy ceramics",
        "texts": [
            " The resultant porosity acts as flaws, which produce stress concentration and result in failure at a lower applied stress according to the following relation: \u03c3f \u00bc \u03c3Sexp nfPtrue\u00f0 \u00de \u00f05:4\u00de where \u03c3f is the flexural strength, \u03c3S is the ceramic material strength without porosity, and nf is an empirical constant. Due to the ionic/covalent bonding and large lattice spacing of ceramics when compared to metals and alloys, slip is relatively difficult, contributing to their brittle modes of failure. The fracture surfaces of ceramics can be divided into four regions, as shown in Fig. 5.1: (1) the fracture source/initiation site, with a radius denoted as af; (2) a smooth mirror region with a highly reflective surface, (3) a misty region that 122 5 Ceramics contains small radical ridges and microcrack distributions, and (4) a hackle region that contains larger secondary cracks. The hackle region may also be bounded by branch cracks in some ceramic systems. The dimensions of the mirror, mist, and hackle regions have been shown to be related to the fracture stress. In highly brittle solids with strong covalent/ionic bonds with little point movements of defects, any issues such as pores, inclusions, or gas bubble entrapments can act as areas for the development of cracks",
            " While the cause for macroscopic fatigue in ductile solids comes from cyclic slip, in ceramics, the main causes for crack growth involve degrading bridging zones behind the crack tip (Fig. 5.2), microcracking, martensitic transformations, and interfacial sliding. 5.2 General Characteristics 123 Many ceramic materials show stable cracking under static or quasistatic stress. In these conditions, one can figure the time rate of crack growth, daf/dt, where af is the radius of fracture source as shown in Fig. 5.1, as a function of the linear elastic stress intensity factor, KE, with a unit of MPa m1/2. The factor KE predicts the stress intensity near the tip of a crack caused by a remote load or residual stress, which has a unit of stress times the root of a distance. KE can be approximated by daf=dt \u00bc AcKE pc , \u00f05:5\u00de where Ac and pc are material constants that are obtained from crack growth experiments under static loading. In general, pc is found to lie between 2 and 4. Equation 5.5 is known as the Paris-Erdogan law"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001352_s1028335815030106-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001352_s1028335815030106-Figure1-1.png",
        "caption": "Fig. 1. Mechanical system.",
        "texts": [
            "SSN 1028 3358, Doklady Physics, 2015, Vol. 60, No. 3, pp. 140\u2013144. \u00a9 Pleiades Publishing, Ltd., 2015. Original Russian Text \u00a9 S.P. Karmanov, F.L. Chernousko, 2015, published in Doklady Akademii Nauk, 2015, Vol. 461, No. 3, pp. 286\u2013290. 140 THE MECHANICAL SYSTEM For modeling the process of displacement of a row ing boat, we consider a mechanical system consisting of the the main body body of mass m and two identical bodies (oars) A1 and A2 attached to it, the mass of which is negligibly small (Fig. 1). The oars make lon gitudinal forward movements with respect to the main body. The motion occurs in a horizontal plane. We introduce the system of coordinates Cxy con nected to the main body. It is assumed that the axis Cx is the body symmetry axis. The oars A1 and A2 are always symmetric to each other with respect to the axis Cx, and their displacement along this axis with respect to the main body is designated as \u03be. Let the main body move translationally along axis Cx at the initial moment of time t = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002651_cmb.2016.0021-Figure9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002651_cmb.2016.0021-Figure9-1.png",
        "caption": "FIG. 9. Third division of the cell in the first quadrant.",
        "texts": [
            " The boundary here being a union of three curves is different from the line boundary (y-axis), but the shape of the outer boundary is still an arc. This process can be deduced theoretically as in Appendix A. The detailed division data can be found in Table 2 in Appendix C. The third division of the cell in the first quadrant begins with its central point of the outer boundary at (3.144,1.86759), and its corresponding radius is equal to 2.19012. After calculation, we find that the shortest length of the division curve of the cell is 2.16159, specifically in Figure 9 and Table 3. Next to case 3, the cell on the right of the division membrane w1 continuously expands in a normal way as the growth of cells in the second quadrant, and the left cell is in the constrained division. When the area of the left cell reaches 2p, the circle point P(3.97075, 0.160892) of the cell locates on the ray of the last cell division line, and the radius of the cell\u2019s outer membrane is 1.97731. Note that, at this moment, the outer boundary, namely the cell\u2019s outer membrane, only intersects with the boundary of the \u2018\u2018dead\u2019\u2019 cell in the fourth quadrant at point A",
            " By calculation, we can find that the length of division line l41 is 2.19165, and there\u2019s no solution in l42. See Table 4 for specific data. After case 4, the division membrane l41 coincides with the membrane of the mother cell. In this way the original cell (the right cell in Fig. 10) has already been an \u2018\u2018inner cell,\u2019\u2019 and it will not grow and divide. Only the cell on the outside will continue to grow (see Fig. 11). The detailed division data can be found in Table 5 in Appendix C. After case 3, the cell on the right of the division membrane w1 in Figure 9 begins to grow; the left cell grows very slowly because it is pressed by its left \u2018\u2018inner cell.\u2019\u2019 This division process is similar with case 1 (Fig. 7). The central point of the outer membrane for the lower right new cell locates at (3.144, 1.86759), and its radius is 3.06062 (Fig. 12). It can be found that the difference between this kind of division and case shows that this division membrane (l63 being the shortest in Fig. 12) bends down, rather than the reverse situation in case 1. See Table 6 in Appendix C for detailed division data"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000645_jae-141897-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000645_jae-141897-Figure5-1.png",
        "caption": "Fig. 5. Magnetic-flux density distribution.",
        "texts": [
            " The limit torque (slip torque) should be larger than the rated torque. The slip torque is determined by the magnetic materials such as the permanent magnets and laminated silicon steel sheets. On the other hand, the applied current has almost no effect on slip torque. To evaluate slip torque, finite element analysis was conducted. The slip torque waveform is shown in Fig. 4. When the HSR is at 5.625 deg, the torque ratio between the LSR, HSR and the stator is approximately 8.5: \u22121: \u22127.5 = Gr: \u22121: \u2212Gr+1. Figure 5 shows a contour plot of the magnetic-flux density distribution. Figure 5(a) shows the MG Generator in its state of equilibrium where HSR is at 0 deg. At position (i), the magnetic flux lines are horizontal. On the other hand, Fig. 5(b) is the slip torque transmitting position. At positions (ii) and (iii), the magnetic flux lines have realigned such that a counterclockwise torque is generated on the LSR. From the Fig. 5(a), the LSR and HSR is rotated at the rated operational speed in order to evaluate the EMF waveform. The obtained waveform is shown in Fig. 6. These waveforms only slightly deviate Table 2 Performance of MG generator Rated torque [kNm] 1021 Power [MW] 2.03 Copper loss [kW] 25.7 Core loss [kW] 29.7 Magnet loss [kW] \u2013 Efficiency [%] 97.3 -2 -1 0 1 2 0 60 120 180 240 300 360 EM F [p .u .] Electrical angle [deg] U VW WU UV VW Fig. 6. EMF waveform. 0 500 1000 0 5.625 11.25 16.875 22.5 Tr an sm is si on to rq ue [k N m ] Rotation angle of HSR [deg] LSR HSR HSR: 120.1 kNm LSR: 120.1 kNm Gr8.5=1021 kNm Fig. 7. Transmission torque waveform. from sinusoidal waveforms. The THD of the phase and line voltage is 7.5% and 4.1% respectively. This is because of the rectangular shape of the laminated silicon steel sheets which are attached to the HSR magnets in Fig. 2(b). Since these laminated silicon steel sheets are rectangular blocks and not curved, the magnetic flux of the HSR at the air gap are square waves, which slightly increases the THD. In Fig. 5(b), torque as a magnetic gear and torque as a generator both contribute to the total torque. However, in order to evaluate only the torque as a generator, the two rotors should be rotated according to the gear ratio with the initial angle of the HSR being 0 deg, as shown in Fig. 5(a). The rated current of 6 Arms/mm2, which is almost the same value as conventional machines, was applied to the coils. Figure 7 shows the transmission torque waveform. The HSR torque (120.1 kNm) indicates the torque as a generator. By multiplying the HSR torque with Gr, the LSR torque (1021 kNm), which is the torque as a magnetic gear is obtained. Table 2 shows the performance of MG Generator. When the magnet loss and mechanical loss are ignored, the efficiency is 97.3%. Copper loss and core loss are almost the same value"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002464_978-981-10-1109-2_5-Figure5.6-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002464_978-981-10-1109-2_5-Figure5.6-1.png",
        "caption": "Fig. 5.6 FRBOI : 1 Fully deployed and compressed position. The compression members (cables) are shown in black. 2\u20133 Intermediate positions. 4 Fully stowed position. The volume reduction ratio VRRFRB\u2212OI = 0.557",
        "texts": [
            " Deployment of this dPZ with uniform angle \u03c8 for each fPZM would cause collisions. Therefore, the thresholds t for fPZMs have been assigned similarly to the Fig. 5.4. Thus the first and last units in the sequence (along the t axis) deploy at first and at last, respectively. In principle, the rigid plate units can be folded in a number of ways. The basic folding methods are illustrated with \u201cFoldable Rigid Barrel\u201d (FRB) [8] below. The \u201coutside-in\u201d (OI) folding mentioned above is the most straightforward method. The side panels can be safely folded out, as shown in Fig. 5.6. As Fig. 5.6 indicates, in principle, the linear connections are compressed. In case of super-pressure, it is the result of the tension in cables. If such structure is underpressurized, the cables are redundant and could be removed. Thismakes such solution particularly suitable for underwater (under-pressured) applications. However, the volume reduction ratio (VRR), that is the relationship of the bounding volumes of the module in stowed (VBs) and deployed (VBd) states, ir relatively high (poor). Although in principle the OI deployment is straightforward and intuitive, in case of super-pressured structures it is not practical"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000534_9781118889664.ch25-Figure4-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000534_9781118889664.ch25-Figure4-1.png",
        "caption": "Fig. 4. Three dimensional temperature distributions when the laser beam moves to the center of track at different scanning strategies. (a-1) unidirectional scan the first track, (a-2) unidirectional scan the second track; (b-1) alternative scan the first track, (b-2) alternative scan the second track.",
        "texts": [
            " This is because the heat content of the part increased as more tracks were deposited and consequently less laser power is required for the subsequent tracks. Furthermore, the laser power increased with the increase of the scanning speed. The influences of scanning speed on the variation trend of laser power with track number are almost the same. It can also be seen that the scanning strategy does not have significant influences on the variation trend and values of laser power with track number. Fig. 4 shows the three dimensional transient temperature distributions when the laser beam moves to the center of track. The previous tracks were reheated as the subsequent tracks were deposited. This can be further verified in Fig. 5 which shows the thermal cycles of the midpoints of track 1, 4, 7, and 10 as the function of time for unidirectional scanning at different scanning speeds: 2 mm/s, 5 mm/s, and 10 mm/s. One can see from Fig. 6 that the cooling rate increased with the increase of scanning speed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000043_978-3-030-13273-6_33-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000043_978-3-030-13273-6_33-Figure2-1.png",
        "caption": "Fig. 2. The geometry and kinematic parameters of the A2 circle.",
        "texts": [
            " Rollers are located on the wheel perimeter, set at an angle of a \u00bc p=4 [rad] to the driving wheel axis. Figures 1 and 2 demonstrate the appropriate geometric measurements and the characteristic points of the system. Point S is the center of mass of the frame, and point H is at the midpoint lying on the A1A2 axis. As far as B is concerned, it is a point belonging to the frame equivalent to point H but on the other wheel axis. The b angle is an angle of the temporary frame rotation frame temporary rotation. In Fig. 2, as an example, one has marked the angular velocity vectors for, respectively, the wheel A2, x2 and the roller xr2. It has been assumed that the driving wheel radiuses are equal, therefore R1 \u00bc R2 \u00bc R3 \u00bc R4 \u00bc R. For the description of the kinematics of any given point of the system, it is advantageous to determine the equation of motion. A right-handed rectangular frame of reference is assigned to each of the movable parts of the system, with its origin being the center of mass of a given part. In addition, we assume that the motion of the robot occurs within the xy plane",
            " First it can be determined from the system geometry that the coordinates of points H and A2 in the xy system are related by the following dependency: xA2 \u00bc xH \u00fe l sinb yA2 \u00bc yH l cos b \u00f01\u00de The frame of the mobile wheeled robot is in planar motion, therefore the speed of point A2 can be written as: VA2 \u00bc VH \u00feVA2H \u00f02\u00de The circle with its center point A2 is in compound motion, and therefore we can determine the velocity of this point A2 with the equation VA2b \u00bc VA2u \u00feVA2w \u00f03\u00de where VA2w \u00bc 0 is the velocity in relative motion, VA2u \u00bc VA2 is the velocity in raising motion, and therefore the absolute velocity of point A2, which we will further mark as VA2 . Let us note the velocity of point A2 in an analytical form is VA2 \u00bc _xA2 i\u00fe _yA2 j \u00f04\u00de and we will determine the velocity of point M shown in Fig. 2a from the relation VM \u00bc VA \u00feVMA \u00f05\u00de where the velocity of point A is VA \u00bc VA2 \u00feVAA2 \u00f06\u00de We can then determine the VAA2 relative velocity vector as VAA2 \u00bc x R \u00bc i j k xx xy xz 0 0 R 2 4 3 5 \u00f07\u00de The following relation stems from the outline of the angular velocity vectors of circle A2, x2 and the frame angular velocity xr: x \u00bc x2 \u00fexr \u00f08\u00de where x2 \u00bc _u2 j1 xr \u00bc _b k1 . The markings appearing in Eq. (8) are, respectively _u2 - the value of the A2 circle angular velocity, _b - the value of the frame angular velocity, i1; j1; k1 - the versors of the system related to the robot\u2019s frame"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001484_s00170-014-6125-8-Figure5-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001484_s00170-014-6125-8-Figure5-1.png",
        "caption": "Fig. 5 Dynamicsmodeling of the manipulation of cylindrical nanoparticles",
        "texts": [
            " At this point, by using the same approach used in deriving the dynamic equations for the displacement of spherical nanoparticles, which was presented in [7, 8], the equations of motion are extracted in the 3D state for cylindrical nanoparticles in liquid medium. It should be mentioned that in the derived equations, the kinematics has been presented by Eqs. (8) through (10) and the internal forces have been presented by Eqs. (13) through (17). The forces applied on the probe tip are derived threedimensionally according to the following equations, and the resultant of these three forces, which is applied on the particle by the probe tip, is calculated (based on Fig. 5). FX \u00bc Fx \u00fe Vcos \u03b8\u00fe max \u00f018\u00de FY \u00bc I \u03b3\u0308 \u00feM \u03b8y H ! cos \u03b3 \u2212 sin \u03b3 cos \u03b3 Fz \u00fe V sin \u03b8\u00fe maz\u00f0 \u00de \u2212 sin2\u03b3 may \u2212 Fy \u00f019\u00de FZ \u00bc I\u03b8 ::\u00feM \u03b8x H sin \u03b8 \u2212 sin \u03b8 cos \u03b8 max \u00fe Fx\u00f0 \u00de \u00fe cos2\u03b8 Fz \u00fe wsin\u03b3 \u00fe maz\u00f0 \u00de \u00f020\u00de V \u00bc FZ \u2212 Fz \u2212maz \u2212 wsin\u03b3 sin\u03b8 \u00f021\u00de w \u00bc FZ \u2212 Fz \u2212 V sin \u03b8 \u2212maz sin\u03b3 \u00f022\u00de FT \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi FX 2 \u00fe FY 2 \u00fe FZ 2 p \u00f023\u00de The presented (Fig. 6) algorithm shows the simulation steps and the required parameters for determining the dynamics of cylindrical nanoparticles in liquid environment"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003280_j.ifacol.2016.11.155-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003280_j.ifacol.2016.11.155-Figure2-1.png",
        "caption": "Figure 2. Body fixed coordinate system (Nelson 1998)",
        "texts": [
            " The method chosen for this research is the analytical as described in (Roskam 2001; Nelson 1998). This was chosen having into account that the aircraft is still under construction while this system is being developed. The variables convention used to de movement description in this paper is the same as used in (Nelson 1998) and a summary of it is shown in Table 1. The aircraft\u2019s modeling can be divided in Longitudinal and Lateral, due to the decoupling of these two kinds of movements and can be developed using as reference the gravity center of the vehicle as shown in Figure 2. This approach has some advantages as it allows several simplifications due to the symmetry of the plane as well as to the rotation axis positions. After the modeling in this reference system, this can be transferred to an inertial reference in a fixed point using (1). [ ?\u0307?\ud835\udc65 ?\u0307?\ud835\udc66 ?\u0307?\ud835\udc67 ] = [ \ud835\udc36\ud835\udc36\ud835\udf03\ud835\udf03\ud835\udc36\ud835\udc36\u03a8 \ud835\udc46\ud835\udc46\u03a6\ud835\udc46\ud835\udc46\ud835\udf03\ud835\udf03\ud835\udc36\ud835\udc36\u03a8 \u2212 \ud835\udc36\ud835\udc36\u03a6\ud835\udc46\ud835\udc46\u03a8 \ud835\udc36\ud835\udc36\u03a6\ud835\udc46\ud835\udc46\ud835\udf03\ud835\udf03\ud835\udc36\ud835\udc36\u03a8 + \ud835\udc46\ud835\udc46\u03a6\ud835\udc46\ud835\udc46\u03a8 \ud835\udc36\ud835\udc36\ud835\udf03\ud835\udf03\ud835\udc46\ud835\udc46\u03a8 \ud835\udc46\ud835\udc46\u03a6\ud835\udc46\ud835\udc46\ud835\udf03\ud835\udf03\ud835\udc46\ud835\udc46\u03a8 + \ud835\udc36\ud835\udc36\u03a6\ud835\udc36\ud835\udc36\u03a8 \ud835\udc36\ud835\udc36\u03a6\ud835\udc46\ud835\udc46\ud835\udf03\ud835\udf03\ud835\udc46\ud835\udc46\u03a8 \u2212 \ud835\udc46\ud835\udc46\u03a6\ud835\udc36\ud835\udc36\u03a8 \ud835\udc46\ud835\udc46\ud835\udf03\ud835\udf03 \ud835\udc46\ud835\udc46\u03a6\ud835\udc36\ud835\udc36\ud835\udf03\ud835\udf03 \ud835\udc36\ud835\udc36\u03a6\ud835\udc36\ud835\udc36\ud835\udf03\ud835\udf03 ] [ \ud835\udc62\ud835\udc62 \ud835\udc63\ud835\udc63 \ud835\udc64\ud835\udc64 ] (1) Being ?\u0307?\ud835\udc65, ?\u0307?\ud835\udc66, ?\u0307?\ud835\udc67 the aircraft velocities with respect to the inertial coordinate system, \u03a6, \ud835\udf03\ud835\udf03 and \u03a8 being respectively the roll angle, the pitch angle and the yaw angle and using \ud835\udc36\ud835\udc36\ud835\udf03\ud835\udf03 = cos (\ud835\udf03\ud835\udf03), \ud835\udc46\ud835\udc46\ud835\udf03\ud835\udf03 = sin (\ud835\udf03\ud835\udf03) and so on"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003566_1.5063231-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003566_1.5063231-Figure2-1.png",
        "caption": "Figure 2. Hot stamped parts in a typical mid-range car",
        "texts": [
            " In 2004 the estimated total consumption of hot stamping steels was considerable: about 60,000 to 80,000 tons in Europe. Annual consumption in Europe during the period 2008-2009 was about 300 000 tons. Japanese and North American car makers are expected to follow this trend, the utilization of hot stamping process will expand rapidly. The main components where this process is utilized in the automotive sector are located in the chassis of a car body, such as A-pillar, B-pillar, bumper, roof rail and tunnel (Figure 2) /2/ Two variants of the press hardening methods exits: direct press hardening and indirect hardening. In the direct process, the blank is heated and formed simultaneously. In the indirect process, the blank is formed first and the formed blank is then heated and end formed using another tool (Figure 3). processes /1/ Typically the steel used in the press hardening process is boron alloyed carbon-manganese steel (22MnB5). The typical chemical composition of 22MnB5 steel is presented in Table 1. The processing of the hardened 22MnB5 steel to its final fully martensitic structure is a result of the following procedure: an annealed material is heated up to its austenitization temperature at around 880-950\u00b0C and to achieve a fully martensitic structure to this material, the cooling rate must exceed 25-30\u00b0C/s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001745_omae2015-41955-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001745_omae2015-41955-Figure1-1.png",
        "caption": "Figure 1. CROSS SECTION OF A PERMANENT MAGNET MOTOR.",
        "texts": [
            " For a harsh Atlantic offshore environment such as in Newfoundland and Labrador, the total cost to replace or repair an ESP unit is very high. Thus, any improvements in ESP reliability and performance can be significant for the offshore oil and gas industries, and the requirements for reducing power consumption and efficient energy usage have become major issues. In recent years, permanent magnet (PM) motor drives have been introduced into electric submersible pump systems by different ESP manufacturers [2-10]. Fig. 1 shows the cross section of a PM submersible motor. A PM motor has a solid rotor made of low loss laminated steel. high energy density rare earth Neodymium Boron Iron (Nd-B-Fe) permanent magnets are embedded inside the rotor for maximum reluctance torque as well as high electric motor torque within minimum volume and weight of the rotor. PM motor drives have the following advantages over the standard induction motor drives,  synchronous operation \u2013no energy wasting slip loss.  40% shorter motor length and 40% lighter motor weight"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000140_978-3-030-43881-4_10-Figure10.46-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000140_978-3-030-43881-4_10-Figure10.46-1.png",
        "caption": "Fig. 10.46 Operational B\u2013H loop of a conventional transformer",
        "texts": [
            " Magnetization of the core is modeled by the magnetizing inductance LM . The magnetizing current iM(t) is related to the core magnetic field H(t) according to Ampere\u2019s law H(t) = niM(t) m (10.104) However, iM(t) is not a direct function of the winding currents i1(t) or i2(t). Rather, the magnetizing current is dependent on the applied winding voltage waveform v1(t). Specifically, the maximum ac flux density is directly proportional to the applied volt-seconds \u03bb1. A typical B\u2013H loop for this application is illustrated in Fig. 10.46. In the transformer application, core loss and proximity losses are usually significant. Typically the maximum flux density is limited by core loss rather than by saturation. A highfrequency material having low core loss is employed; in a transformer-isolated switching converter, ferrite typically is used. Both core and copper losses must be accounted for in the design of the transformer. The design must also incorporate multiple windings. Transformer design with flux density optimized for minimum total loss is described in Chap"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001822_jae-150004-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001822_jae-150004-Figure1-1.png",
        "caption": "Fig. 1. Healthy and eccentric submersible motor considering slotting effect. O1 is the center of the stator and O2 is the center of the rotor. In a healthy motor, O1 and O2 coincide. In case of the static eccentricity, O2 is the center of rotation. In the case of the dynamic eccentricity, O1 is the center of rotation. In the case of the mixed eccentricity, the center of rotation can be anywhere between O1 and O2.",
        "texts": [
            " In this paper, a direct analysis of the magnetic flux density of air gap considering slotting effect is presented in the submersible motor. The type of air-gap eccentricity fault can be distinguished directly by judging whether there are characteristic frequencies in the induced voltage spectra of search coil, which is positioned around the stator tooth. Firstly, based on the MMF and permeance wave approach, the air-gap length should be obtained under eccentricity fault considering slotting effect. The eccentricity can be easily visualized in the submersible motor following Fig. 1. The air-gap permeance considering slotting effect has been described in the healthy motor in [24]. Based on it, the origin of the stator frame of reference is placed on the center axis of the stator slot. If the slotting effect is considered under eccentricity fault, the air-gap length is given by g(\u03d5, t) = g0 + gs cos\u03d5+ gd cos(\u03d5\u2212 \u03c9rt) + \u03bd=\u221e\u2211 \u03bd=1 a\u03bd cos \u03bdZ1\u03d5+ \u03bd=\u221e\u2211 \u03bd=1 b\u03bd cos \u03bdZ2(\u03d5\u2212 \u03c9rt) (3) where g0 is the air gap constant, gs is the magnitude of static eccentricity, gd is the magnitude of dynamic eccentricity, \u03d5 is the angular position in the stator frame of reference, Z1 is the number of the stator slots, Z2 is the number of the rotor slots, \u03c9r is the rotational velocity of the rotor in radians per second which is governed by the slip s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001285_esda2014-20575-Figure2-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001285_esda2014-20575-Figure2-1.png",
        "caption": "Figure 2. Gears mesh (a) Mesh of the wheels; (b) mesh of the teeth.",
        "texts": [
            " In our simulations, only a couple of teeth were in contact and the two standard involute wheels used were identical. Their characteristics can be found in Table 1. A torque of 150N.m was applied to the wheels, for a rotational speed of 0.6rad.s -1 . The duration of the simulation was 0.1s and the teeth were in contact on a distance of approximately 1mm. A commercial finite element program (Abaqus, Dassault Syst\u00e8me) was used in which the size of the quadratic elements of the mesh varied from 1.25\u00b5m in the contact zone to 20\u00b5m in the farthest parts (Figure 2). Seventeen points along the contact path were used in order to calculate the associated stiffness, while ten were used for the bending stiffness. The contact between elements is modeled with a constant friction coefficient of 0.2 and is not lubricated. Negative surface replicas made of a silicon rubber material (Struers, Repliset F5) were used to assess the 3D texture of teeth surfaces finished respectively by grinding and powerhoning finishing process. Topographical features of replica surfaces were measured in three locations by a threedimensional white light interferometer, WYKO 3300 NT (WLI) (Figure 3)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002736_j.mechmachtheory.2016.07.025-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002736_j.mechmachtheory.2016.07.025-Figure1-1.png",
        "caption": "Fig. 1. Geometry of enveloping curves. (a) Generating and envelope curves and trace of contact point along the curves, (b) auxiliary rigid body Tnp for kinematic analysis.",
        "texts": [
            " Griffis [9] introduced higher-order reciprocity for two surfaces in point contact where a set of contact kinematics equations are derived for higherorder conjugate motion using twist and twist derivative in the dual vectors set up. In this paper, we use the framework and the contact kinematics equations of Griffis for our analysis. The basic geometric entities in envelope theory are two curves in a point contact which are in relative roll-slide motion. Without loss of generality, one of them can be assumed to be fixed. The generating curve and its envelope are labeled as the mcurve and f-curve respectively; their point of contact is C (Fig. 1 (a)). We refer to the contact normal line as the n-line (labeled n in Fig. 1 (b)) in this paper for brevity wherein n is assumed to be oriented and its orientation is away from the f-curve, or towards the m-curve. A right-handed coordinate system, XY Z, is attached to the fixed frame with its origin at the point of contact C. The positive Y-axis is chosen along the n-line which in the direction away from the material side of the f-curve at C and the positive X-axis is chosen such that positive Z-axis points out of the plane of the paper. Let the centers of curvature of the f-curve and mcurve be Cf and Cm with coordinates Cf = [0, 1/kf]T and Cm = [0, 1/km]T respectively, where km and kf are the signed curvatures of the m-curve and f-curve at C. It may be noted that (km \u2212 kf) > 0 for all contacting geometries in the chosen convention. Let T = T + 40 be the common tangent line at C directed along the instantaneous velocity of the point of the m-curve which coincides with C. Also, n = n + 40 be the contact normal line (Fig. 1 (b)). Then, the line p = p + 40 which passes through C is such that Tnp is a right handed mutually orthogonal triad of lines. T , n, and p are unit vectors. A virtual three-link mechanism consisting of the f-curve(body-1 in [9]), the m-curve (body-2 in [9]) and Tnp based virtual body (body-3 in [9]) is used to derive the contact kinematics between the two curves. The instantaneous relative velocity between the points of the curves at the point of contact C is directed along T. Let s\u0307m and s\u0307f be the speeds with which the point of contact moves along the m-curve and f-curve respectively and vc be the speed of the point of the m-curve which coincides with C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001996_sii.2015.7404997-Figure12-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001996_sii.2015.7404997-Figure12-1.png",
        "caption": "Fig. 12. Initial angle of 3-link manipulator is set to q = [\u22120.05,\u2212\u03c0/2+ 0.05,\u2212\u03c0/2 + 0.05]T as (a). As a result of free response(fi = 0), manipulator converged at q = [\u2212\u03c0, \u03c0/2, ]T as (b).",
        "texts": [
            "(38)\u223cEq.(41) is 0f\u03021 = [0, 0.002, 0.003]T, 0f\u03022 = [0, 0.006, 0.003]T. And translational acceleration of the center of mass of link 1 and link 2 will become 0f\u03021/m1 = [0, 0.002, 0.003]T, 0f\u03022/m2 = [0, 0.006, 0.003]T, there is no contradiction between the relationship of acting force and acceleration at the time of 2\u00a9 similar to time of 1\u00a9. Then, we examined the acting force and constraint force acting between each link of 3-link manipulator. We set the posture (q = [\u22120.05,\u2212\u03c0/2 + 0.05,\u2212\u03c0/2 + 0.05]T) in Fig.12(a) that was displaced by 0.05 angle from the balance point of q = [0,\u2212\u03c0/2,\u2212\u03c0/2]T as the initial posture, finally, we made the free response simulation that converge to balance point of Fig.12(b). Here, Fig.13 shows screen shot of simulation and Fig.14 shows the time response of the y-component of the force 0f1 = [0f1x, 0f1y, 0f1z]T 0f2 = [0f2x, 0f2y, 0f2z]T, 0f3 = [0f3x, 0f3y, 0f3z]T, acting on each joint and the frictional force ft acting on hand. Fig.15 shows the time response of the z-component and constraint force fn,and Table 3 shows the physical parameters, besides Table 4 shows the initial value and final value. In Fig.14 and 15, we set the time of t = 3.01 as shows the direction of acceleration and the force acting on direction y, z"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000052_978-3-030-24237-4-Figure10.9-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000052_978-3-030-24237-4-Figure10.9-1.png",
        "caption": "Fig. 10.9 Configuration of (a) facing and (b) and slot cutting",
        "texts": [],
        "surrounding_texts": [
            "The multiple (n) teeth on the milling head with a diameter of D rotating with a speed of N in rpm perform multiple cuts over the material surface with a thickness of B during the milling operation. As illustrated in Fig. 10.10 (left), the direction of the rotational axis is perpendicular to the feed direction, and these two axes are parallel to the material surface. The material surface area being removed is defined by the length of cut L and the width of cut (i.e., the tooth length) as indicated in Fig. 10.10 (right). The cutting speed v is then determined by Eq. 10.1. Meanwhile, the material sample mounted on a movable stage moves toward the milling head with a feed rate f (millimeter per minute), and the feed distance per single tooth cut (Lfeed_tooth) is f/N/ n (or the feed distance per single revolution Lfeed \u00bc f/N). 278 10 Design for Manufacturing The cutting time Tcut is calculated by Eq. 10.2, whose length of initial offset position (LA) can be further estimated by this simple proof provided below1: LA \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi t D t\u00f0 \u00de p \u00f010:6\u00de where t is the width/depth of cut. On the other hand, the MRR with LA L is MRR LWB Tcut \u00bc fWB \u00bc kv \u03c0D WBLfeed \u00bc WBNnLfeed tooth \u00f010:7\u00de"
        ]
    },
    {
        "image_filename": "designv11_64_0002354_tec.2015.2490801-Figure8-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002354_tec.2015.2490801-Figure8-1.png",
        "caption": "Fig. 8. Quarter of a 2-D model of the permanent magnet motor.",
        "texts": [
            " This is consistent294 with the results presented in [20], where forces obtained with295 the MST and ESM (using external field) are compared with dif-296 ferent mesh densities (the sum of stator field components is the297 external field for the rotor).298 It is evident from Fig. 7 that a major part of the torque ripple299 produced by the squirrel-cage induction motor is due to stator300 iron to rotor iron interaction. In this example, there is also a301 notable phase shift between the stator iron to rotor iron torque 302 component and the other torque components. 303 B. Permanent Magnet Motor 304 Fig. 8 presents the model geometry of the permanent magnet 305 motor and Table II the essential motor parameters. The rotor has 306 four surface-mounted magnets with 1-T remanence flux density 307 and 1.05 \u03bc0 permeability. The BH-curve of the iron cores and 308 shaft are presented in Fig. 3. The iron cores are assumed non- 309 conducting, whereas the shaft and magnets have a conductivity 310 of 4.3e6 S/m and 6.7e5 S/m, respectively. 311 The permanent magnet motor fields are decomposed into four 312 components: the stator iron field, the stator coil current field, the 313 rotor iron field, and the rotor magnets field"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0001275_s11431-014-5703-1-Figure1-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0001275_s11431-014-5703-1-Figure1-1.png",
        "caption": "Figure 1 Schematic of 3BSD nozzles. (a) With position of N=0\u00ba, Ny=0\u00ba; (b) with position of N=90\u00ba, Ny=0\u00ba.",
        "texts": [
            " The rest of this paper is organized as follows. Problem formulation is given in Section 2. A second-order characteristic model is built for the revolute pairs in Section 3. The control strategy is presented in Section 4. Experimental results are provided and analyzed in Section 5. Finally, a brief conclusion is drawn in Section 6. The fundamental and the kinematic model of 3BSD nozzles have been discussed in ref. [8]. For the convenience of readers, some results will be presented in this section. Figure 1 shows the schematic of 3BSD nozzles, where 1 denotes the engine; 2, 3, and 4 denote the first, second, and third duct, respectively; 5, 6, and 7 denote the first, second, and third bearing, respectively;  denotes the inclination angle of the end sections of each duct; 1, 2, and 3 denote the rotation angle of the adjacent ducts, respectively; TN denotes the thrust vector force of the nozzle; and Obxbybzb denotes the body axis of the nozzle. For simplicity, the direction of TN is assumed to be consistent with the direction of the 3BSD nozzle's outlet"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0002344_ilt-08-2015-0119-Figure13-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0002344_ilt-08-2015-0119-Figure13-1.png",
        "caption": "Figure 13: Results of QFH calculation for dyn. peak load of 30 MPa",
        "texts": [
            " Asperity contact is restricted to a small portion of the surface and only to an area behind the crack relative to the direction of rotation. Thus, a material overload and not asperity contact seems to be the cause for bearing damage at 36 MPa. Figure 12: Journal bearing with fatigue damage, material: SnSb12Cu6ZnAg, spec. dyn. peak load: 36 MPa The QFH approach enables an evaluation of the loading while considering the orientation dependent fatigue parameters of the material. According to eq. (8), the distribution of the failure value F on the developed sliding surface of the lower half of the bearing is shown in Figure 13 and Figure 14. In d L ub ri ca tio n an d T ri bo lo gy 2 01 6. 68 . Figure 14: Results of QFH calculation for dyn. peak load of 36 MPa For 30 MPa the maximum failure value F of 0.97 is reached at a circumferential angle of 107\u00b0 and in the axial middle of the sliding surface. For 36 MPa the maximum failure value F of 1.19 is reached at a circumferential angle of 103.5\u00b0 and (also) in the axial middle of the sliding surface. As it is assumed that a value F \u2265 1 leads to damage in the continuum, it can be shown, that for 30 MPa no bearing failure is expected since F does not exceed a value of 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0003081_urai.2016.7625770-FigureI-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0003081_urai.2016.7625770-FigureI-1.png",
        "caption": "Fig. I, Continuous measurement of the position coordinates.",
        "texts": [
            " Although the method is made to compensate for the zero-offset, the kinematic error due to uncertainties in the D-H parameters can also be corrected to some extent since the zero-offset compensation takes into account the total error. To measure the endpoints, we first attach a bar to the end effector to be perpendicular to the axis of rotation. Then, we fix the laser tracker target at the end of the bar. The length of the bar is to be more than lOcm. All the joints are fixed at the reference angle, except one joint that is to be rotated. We start from the first joint and then goes to outward joints one after another. Refer to Fig.I. Step 1: The joint of rotation is made to gradually rotate by 2 to 5 degrees at each instance. Step 2: The remaining joints maintain their fixed angles. Step 3: Measure the endpoint coordinates using the laser tracker. Step 4: Save the angles of all the joints together with the laser tracker measurement. Step 5: Measure to the maximum travel angle of the joint of rotation. Step 6: After finishing measuring the joint of rotation, the angle of the joint is made to come back to a fixed nominal angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000665_icinfa.2015.7279796-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000665_icinfa.2015.7279796-Figure3-1.png",
        "caption": "Figure 3. Illustration of the projection operator",
        "texts": [
            " Definition 1: Consider a convex compact set with a smooth boundary given by { / ( ) }, 0 1n c R f c c\u03b8 \u03b8\u2126 = \u2208 \u2264 \u2264 \u2264 (7) where : nf R R\u2192 is the following smooth convex function 2 max 2 max ( 1)( ) T f \u03b8 \u03b8 \u03b5 \u03b8 \u03b8 \u03b8 \u03b8 \u03b5 \u03b8 + \u2212 = . (8) Here, max\u03b8 is the norm bound imposed on the vector \u03b8 , and 0\u03b8\u03b5 > is arbitrarily chosen. The projection operator is defined as , ( ) 0 Pr ( , ) , ( ) 0 0 , ( ) , ( ) 0 0 T T Y if f oj Y Y if f and f Y f fY Y f f f if f and f Y \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u23a7\u23aa\u23aa\u23aa <\u23aa\u23aa\u23aa\u23aa= \u2265 \u2207 \u2264\u23a8\u23aa\u23aa\u23aa \u2207 \u2207\u23aa \u2212\u23aa\u23aa \u2207 \u2207\u23aa\u23a9 \u2265 \u2207 > (9) The projection operator is shown in Figure 3. Control law: \u02c6( ) ( )( ( ) ( ))gu s C s s k r s\u03b7= \u2212 \u2212 (10) where ( )r s and \u02c6( )s\u03b7 are the Laplace transforms of ( )r t and \u02c6\u02c6( ) ( ) ( )Tt t X t\u03b7 \u03b8= , 11 / ( )T g lk c A B\u2212=\u2212 , ( )C s is a BIBO-stable and strictly proper transfer function with DC gain (0) 1C = . The variable 0k > is a feedback gain and ( )D s is any transfer function that leads to the strictly proper stable ( )( ) 1 ( ) wkD sC s wkD s = + (11) The low-pass filter manages to decouple the adaptation and robustness. The choice of proper channels may control the frequency of signals and filter the high-frequency signals in control inputs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_64_0000803_ccece.2014.6901107-Figure3-1.png",
        "original_path": "designv11-64/openalex_figure/designv11_64_0000803_ccece.2014.6901107-Figure3-1.png",
        "caption": "Fig. 3. Elliptical model of a hysteresis loop",
        "texts": [
            " Br is called the remanence which is the residual value of flux density when the applied field becomes zero. Hc is called the coercive force or coercivity of the material which is the negative value of the applied magnetic field intensity, required to force the flux density to zero. Elliptical modeling is a way to approximate the shape of the hysteresis loops of a material. In elliptical modeling, the hysteresis curves of the rotor material are approximated by a group of inclined ellipses of similar shapes [4-6]. The elliptical approximations of the hysteresis curves are shown in Fig. 3. The trajectory of a B-H curve lies on the ellipse if the motion around the ellipse is in the counter clockwise direction. The lag angle between B and H remains constant as long as the direction of the motion remains counter clockwise. When the direction of motion changes to clockwise, it results in a movement between the inner ellipses with different lag angles. This phenomena is illustrated in Fig. 3 for one inner ellipse and one outer ellipse with two different lag angles \u03b4 and \ud6ff , respectively. The flux density B and the magnetic field intensity H in an elliptical model can be expressed as follows [12], \ud435 = \ud435 cos (\ud714\ud461 \u2212 \ud713 \u2212 \ud713 ) (1) \ud43b = \ud435 \ud707 \ud450\ud45c\ud460(\ud714\ud461 \u2212 \ud713 \u2212\ud713 + \ud6ff) (2) \ud713 = tan \ud45f sin\ud6ff \ud45d\ud707 \ud459 \ud45d\ud461 \ud707 \ud45f + \ud45f \ud45d\ud707 cos \ud6ff (3) where \ud435 is the maximum flux density of the rotor material, \u03bc is the permeability of the elliptic hysteresis loop, \u03c9 is the synchronous angular frequency, \ud713 (\ud713 = \ud45d\ud703 ; \ud703 is the mechanical angle of the rotor and p is the number of pole pairs) is the electrical angle coordinate in the stator frame, \ud713 is the phase shift and \ud6ff is the hysteresis lag angle between B and H"
        ],
        "surrounding_texts": []
    }
]