[
    {
        "image_filename": "designv11_100_0003658_icsens.2012.6411062-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003658_icsens.2012.6411062-Figure1-1.png",
        "caption": "Figure 1 Specifications of pH electrodes.",
        "texts": [
            " The proposed system includes the pH detecting system, data search system, statistics system, data transmission, web publishing tool. I. EXPERIMENTAL The polyethylene terephthalate (PET) was used as a substrate of hydrogen ion sensor. A 2 in.-diameter and 99.99% purity ruthenium metal target was deposited upon PET substrate using the R.F. sputtering for the RuO2 thin film as the sensing membrane with an area of 4 mm \u00d7 4 mm. The pH electrodes were finished with an area of 20 mm \u00d7 35 mm. According to above-mentioned specifications, the detailed interpretation is exhibited in Fig. 1. [10]. 978-1-4577-1767-3/12/$26.00 \u00a92012 IEEE In this research, the characteristic determination of sensor was analyzed by voltage-time (V-T) measurement system which consists of the front-end of sensing device and the signal treatment. The sensing device including a sensor and a reference electrode (Ag/AgCl) was immersed in the test solution. And environment temperature was controlled by the proportional integral derivative (PID) temperature controller. Then the signal treatment system includes an instrumentation amplifier (LT1167), and one data acquisition card (DAQ card) (Model: NI USB-6210, National Instrument Corp"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001300_s00170-021-07837-2-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001300_s00170-021-07837-2-Figure1-1.png",
        "caption": "Fig. 1. The motorized spindle system",
        "texts": [
            " Section 2 establishes the dynamic model for the shaft-bearingtoolholder structure in a motorized spindle. Section 3 presents a novel method for evaluating the lifetime of the bearing group in a motorized spindle based on the dynamic model. Section 4 experimentally verifies the stiffness calculated by the dynamic model. Section 5 compares the lifetimes of each bearing and bearing group under different conditions of bearing stiffness, clamping force and rotating speed. Section 6 concludes the paper. The internal structure model of a motorized spindle is shown in Fig. 1. Due to the shaft deflection, gyroscopic moment and nonlinearity of bearing stiffness, the bearing load motivated by cutting force is dynamic and nonlinear across different positions. Therefore, the response of bearings in various positions differs. In this case, a simplified dynamic model of a motorized spindle should be established to determine the relationship between cutting forces and bearing loads. On the basis of its internal structure, a motorized spindle can be divided into three parts, namely, shaft, bearing and toolholder"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003445_ht2012-58139-Figure13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003445_ht2012-58139-Figure13-1.png",
        "caption": "Fig. 13: Maximum principal stress distribution on the welded plate.",
        "texts": [],
        "surrounding_texts": [
            "The conclusions are summarized as follows. (1) Weld pool driving forces like surface tension gradient variation significantly alters the temperature distribution in the welded joint. (2) It has been shown as to how incorporation of weld pool dynamics modeling into the residual stress distribution prediction is vital. Only when weld pool dynamics is considered can the accurate thermal distribution be obtained. (3) The developed mathematical model provides the necessary framework to perform coupled welding analysis. (4) The preset approach can prove to significantly better prediction of residual stresses and also develop efficient stress relief mechanisms."
        ]
    },
    {
        "image_filename": "designv11_100_0003846_s1068798x11030129-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003846_s1068798x11030129-Figure1-1.png",
        "caption": "Fig. 1. Involute engagement of gear transition with asym metric tooth profiles.",
        "texts": [
            " A self brak ing transmission with such teeth was designed in [2]. In the present work, we focus on synthesis of the tran sition curves of the tooth tips in such transmissions. Insufficient attention was paid to those curves in [2]. Consider a gear transmission with the same geo metric parameters as in [2]: number of gear teeth z1 = z2 = 6; engagement angles of working and braking pro files \u03b1W = 20\u00b0 and = 80\u00b0, respectively; the interax ial distance aW = 300 mm. The other parameters will be given as we proceed. The basic transmission is shown in Fig. 1. (As yet, the involute curves of the gears do not have transition curves.) Synthesis of the transition curves at the tooth base is divided into two stages. First, we calculate the tran sition curve of the tooth base such that gear rotation is associated with successive contact of the base with the tooth tip of the conjugate gear (as in gear knurling, for example). To this end, we construct curves (roulettes) associated with initial gear circumferences that roll over one another. The roulettes are described by the normals passing though the instantaneous poles of rotation [3]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.4-1.png",
        "caption": "Figure 11.4: Diagonal locations at critical body closure openings.",
        "texts": [
            " This provides the rationale for using diagonal distortions as indicators of vehicle squeak and rattle performance in design iteration studies. The purpose of this study is to assess squeak and rattle performance implications by stiffening the weak body joints using structural foam. Epoxy-based structural foam was employed to stiffen the B-pillar to roof, D-pillar to roof and D-pillar to sill joints as shown in Figure 11.3. The effect of structural foam on body-in-prime is quantified with diagonal distortions at closure openings (Figure 11.4) under static bending and torsional loads. A comparison of these diagonal distortions with and without structural foam is given in Table 11.1. This table indicates that stiffening the three above-mentioned joints with structural foam resulted in 15% Table 11.1: B-I-P stiffness improvement with structural foam Closure openings Normalized diagonal distortions Baseline Foamed body % Improvement Left_frontdoor_top-Aebottom-B 0.715 0.523 26.8 Left_frontdoor_top-Bebottom-A 0.732 0.586 19.9 Left_reardoor_top-Bebottom-C 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.3-1.png",
        "caption": "Fig. 7.3 C-manifolds for RR-type and RP-type 2-link robots",
        "texts": [
            " Therefore, it is clear that the topology of the C-manifold for a dynamic system is determined by its kinematics, while the geometry of the C-manifold is determined by its dynamics. The most simple but typical examples are the two-link robots: one has two revolute joints (RR-type) and the other one is of revolute-prismatic (RP) type, as shown in Figure 7.2. The C-manifold of the RR-type arm is homeomorphic to the surface of 2-torus T 2 S1 \u00d7 S1, while the second one has a C-manifold topologically equivalent to a 2D cylindrical surface S1 \u00d7 I1, as shown in Figure 7.3, where I1 = [0, 1] is the 1-dimensional unity interval. In addition, for each type of the two-link robot arms, a variation of the kinematic parameters, such as the link length and joint off-set, and a variation of the dynamic parameters, such as the mass, mass center, and the moment of inertia, will cause each individual C-manifold to have a different geometry, while the topology remains invariant. In the later sections of this chapter, we will further discuss how to represent both the topology and geometry for a C-manifold"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003156_1.4005598-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003156_1.4005598-Figure1-1.png",
        "caption": "Fig. 1 Structure of a single screw compressor",
        "texts": [
            " A new numerical contour method is proposed to model the groove bottom profile by identifying the contact points on the end face of the cutter to a set position on the spiral groove. Results show that the new method can calculate accurately both the boundary and the inner contact line and thus simulate the groove bottom profile exactly. This method could also be used to simulate other profiles of machines, such as rotor profiles of twin screw compressors and screw pumps. [DOI: 10.1115/1.4005598] Keywords: contour method, envelope method, boundary contact line, groove bottom profile, single screw compressor In a single screw compressor as shown in Fig. 1 [1], sealing performance between the star-wheel tooth and the spiral groove is very important, because it is sensitive to the flowrate of the compressor [2,3]. For this reason, the tooth tip profile is designed according to the profile of the spiral groove bottom to make sure that the tooth tip could hermetically mesh with the groove bottom with enough small clearances between them. When the spiral groove is machined by turning method, circular profile of the tooth tip could be designed [4,5]. In recent years, a milling method is proposed to machine the spiral groove [5]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003416_amm.271-272.1032-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003416_amm.271-272.1032-Figure1-1.png",
        "caption": "Fig. 1 Rotor system with misaligned journal bearings: (a) rotor-bearing system; (b) coordinate system",
        "texts": [
            " Accordingly, this article focuses on the modeling of rotor system supported on a misaligned journal bearing, as well as its nonlinear dynamic characteristics of the rotor-bearing system with the small misalignment and mass unbalance. In this section, the mathematical model of a rotor system supported on misaligned journal bearings and nonlinear oil film forces are mainly developed. Considering the complexity of the rotor-bearing system, hence, some necessary hypotheses are introduced: (a) the structure of rotor-bearing system is symmetrical, (b) both the bearing misalignment and mass unbalance are small, and (c) the long journal bearing theory is cited. Figure 1 illustrates a rotor system supported on three journal bearings, in which \u03b4 is a misalignment of the middle bearing. The purpose of the present paper is to explain the mechanism of vibrations and reveal its dynamic characteristics of the flexible rotor system with a misaligned journal bearing, thereby, for the above symmetrical rotor system, then its motion equations can be expressed )( )( )sin()()( )cos()()( )( )( 2333 2333 2 32122 2 32122 2111 2111 yykFym mgxxkFxm tMayykyykyM tMaMgxxkxxkxM yykFym mgxxkFxm y x y x \u2212\u2212= +\u2212\u2212= \u2126\u2126+\u2212\u2212\u2212\u2212= \u2126\u2126++\u2212\u2212\u2212\u2212= \u2212\u2212= +\u2212\u2212= (1) where F1x, F1y, F3x, F3y are the oil film forces of journal bearings in x and y directions, and other parameters are presented in Fig. 1. In this work, the long bearing theory is introduced, thus, the Reynolds equation for the oil film pressure, p, in Figure 2 yields \u03b8 \u03b8 \u03d5 \u03b8\u00b5\u03b8 cos)2( 2 1 12 1 3 2 e hph R + \u2202 \u2202 \u2212\u2126=      \u2202 \u2202 \u2202 \u2202 (2) where )cos1(cos \u03b8\u03b5\u03b8 +=+= cech . Therefore, the oil film forces in radial and tangent directions are gotten by integrating over the area of the journal sleeve under the boundary condition 00 == == \u03c0\u03b8\u03b8 pp and half-Sommerfeld condition 2 2 1 2 0 2 2 2 3 4 0 2 ( ) cos 6 ( 2 ) 2 ( ) sin 6 ( 2 ) 2 B j j Bjr j j j j j B j j Bjt j j j j j d dR F p R dzd BR E E c dt dt d dR F p R dzd BR E E c dt dt \u03c0 \u03c0 \u03d5 \u03b5 \u03b8 \u03b8 \u03b8 \u00b5 \u03d5 \u03b5 \u03b8 \u03b8 \u03b8 \u00b5 \u2212 \u2212  = = \u2126 \u2212 +         = = \u2126 \u2212 +        \u222b \u222b \u222b \u222b j=1,3 (3) where 2 1 2 2 2 (1 )(2 ) j j j j E \u03b5 \u03b5 \u03b5 = \u2212 + \uff0c 2 2 3/ 2 2 1 8 (1 ) 2 (2 ) j j j E \u03c0 \u03b5 \u03c0 \u03b5   = \u2212  \u2212 +   \uff0c 3 2 1/ 2 2(1 ) (2 ) j j j j E \u03c0\u03b5 \u03b5 \u03b5 = \u2212 + \uff0c 4 2 2 2 (1 )(2 ) j j j j E \u03b5 \u03b5 \u03b5 = \u2212 + ",
            " From equation (3), the oil film forces in x and y directions are obtained    +\u2212= \u2212\u2212= jjtjjrjjjjjy jjtjjrjjjjjx FFyxyxF FFyxyxF \u03d5\u03d5 \u03d5\u03d5 cossin),,,( sincos),,,( (4) where cos j j j x e \u03d5 = , sin j j j y e \u03d5 = , 2 2 j j je x y= + , 2 j j j j j j d y x x y dt e \u03d5 \u2212 = , j j j j j j de x x y y dt e + = . Generally, the oil film forces ),,,( jjjjjx yxyxF , ),,,( jjjjjy yxyxF are the functions of rotor displacements and their velocities, however, if a misalignment 2 2 x y\u03b4 \u03b4 \u03b4= + of the journal bearing exists as shown in Fig.1, the forces in the middle one yield 3 3 3 3 3 3 3 3 3 3 3 3 ( , , , , ) ( , , , , ) x x y y F F x y x y F F x y x y \u03b4 \u03b4 =  = (5) The non-dimensional parameters are introduced on the bearing clearance c and disc weight mg. / , / , / , / , / , / , / , / , / , / , / , , / 1,2,3. 1,3. i i i i x x y y jx jx jy jy jr jr jt jt x x c y y c n M m K kc mg a c c c f F mg f F mg f F mg f F mg t c g i j \u03b1 \u03b4 \u03b4 \u03c4 \u03c9 = = = = = \u2206 = \u2206 = = = = = = \u2126 = \u2126 = = (6) and denote ,/,/ xddxxdtdx \u2032== \u03c4 . Accordingly, equation (1) can be cast into the non-dimensional form, and the six second-order equations can be converted into twelve first-order ones"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003778_978-3-642-17234-2_4-Figure4.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003778_978-3-642-17234-2_4-Figure4.8-1.png",
        "caption": "Fig. 4.8 Geometrical representation of the Euler angles: u, h and w.",
        "texts": [
            " From the practical point of view, the choice of three direction cosines as independent parameters is not convenient. One has to look, then, for other solutions. The most efficient method was devised by Leonhard Euler. He defined a system of three angular parameters, attached to a group of three successive rotations about three conveniently chosen directions. In this way, the transition from S(Oxyz) to S0\u00f0O0x0y0z0\u00de is realized. There are three successive steps in the transition, as follows: (a) A direct (i.e. counterclockwise) rotation of angle u; in the xy-plane about the Oz-axis, until the new axis Op (Fig. 4.8a) is orthogonal to Oz0: (The orientation of S0 relative to S is given!) Thus, we go from the frame S, of unit vectors i, j, k, to the frame Opqz, of unit vectors t1; t2; k (Fig. 4.8b). The transformation relations are: t1 \u00bc i cos u\u00fe j sin u; t2 \u00bc i sin u\u00fe j cos u; k \u00bc k: \u00f04:4:1\u00de The Op axis is known as the line of nodes and u \u2013 as the precession angle, varying from 0 to 2p. (b) A direct rotation of angle h, in the Oqz-plane, about the line of nodes Op, until Oz coincides with Oz0; i.e. the transition from Opqz to Oprz0; of unit vectors t1; t3; k0: The transformation formulas are: t1 \u00bc t1; t3 \u00bc t2 cos h\u00fe k sin h; k0 \u00bc t2 sin h\u00fe k cos h: \u00f04:4:2\u00de The angle h is called angle of nutation and takes values from 0 to p: (c) A direct rotation of angle w, in the Opr plane, about the Oz0-axis, until Op coincides with Ox0: If i0; j0; k0 are the unit vectors of S0; then the transition from Oprz0 to Ox0y0z0 is given by: i0 \u00bc t1 cos w\u00fe t3 sin w; j0 \u00bc t1 sin w\u00fe t3 cos w; k0 \u00bc k0: \u00f04:4:3\u00de The angle w is the angle of self-rotation and takes values from 0 to 2p: The angles u; h;w are called Euler\u2019s angles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001253_robot.2005.1570279-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001253_robot.2005.1570279-Figure5-1.png",
        "caption": "Fig. 5. Redundancy of the three-fingered gripper used as an example.",
        "texts": [
            " Such solution is selected for the application example presented in the next section: a full search is performed in the upper tree levels (those corresponding to the finger assignments) thus resulting in a set of leaf nodes; and then heuristic searches are performed in the lower tree levels, starting at each previous leaf node. As an example, the proposed grasp redundancy resolution method will be used to select the optimum configuration in a 2D planar problem. A three-fingered articulated gripper like the one shown in Fig. 5 is selected. The gripper has 4 DOF: 3 of them translational (opening and closing of the fingers) and one rotational (abduction of the two articulated fingers, which are coupled). As it can be seen in the figure, multiple different grasps can be performed with such gripper, given three contact points on the object surface. First, three different finger assignments can be chosen (rows of the figure) and then, for each assignment it is possible to reach the contact points with different abduction angles (columns of the figure)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002526_kem.572.397-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002526_kem.572.397-Figure3-1.png",
        "caption": "Fig. 3 The direct injection nozzle Fig. 4 The conical nozzle with swirl chamber",
        "texts": [
            " The width of the upper one made of ZCuPb10Sn10 with hardness HBS70, is 4mm. Structure of the specimen is shown as Fig.2. Lubrication Method. The oil-air lubrication is adopted. The air pressure is 0-0.5Mpa. The L-TSA46 machine lubricating oil supply volume controlled exactly by adjusting the time of oil supply is 0-50ml/h. Lubricating oil and air are transported from oil-air pipe through the nozzle to the lube point. Oil-air nozzles are direct injection nozzle without swirl chamber and conical nozzle with swirl chamber shown as Fig.3 and Fig.4. The rotation speed of the test is 210rpm with temperature of laboratory 17.5 \u2103. The load from 50N gradually increases to 1500N with the unit of 300N for every 20 minutes in order to ensure the accuracy of the test temperature and the friction coefficient. Oil-air lubrication test at both direct nozzle and conical nozzle with different oil supply (3ml/h, 10ml/h, 15ml/h, 20ml/h, 30ml/h, 40ml/h, 50ml/h) and air supply 2.25m 3 /h was carried out. Oil-air lubrication test at direct nozzle with different air supply (1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002905_175355510x12810198707653-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002905_175355510x12810198707653-Figure2-1.png",
        "caption": "Figure 2 shows sketch maps of different laser waves [continuous wave (CW), sine wave and square wave]. The other basic process parameters are expressed as follows: laser beam diameter is 0?5 mm; scanning speed is 6 mm s21; powder feeding rate is 8 g min21; gas flowrate is 8 L min21; utilisation rate of laser energy is 30%. In simulation process, both the substrate and metal powder are stainless steel AISI 316L, and thermal physical parameters of the material change with the change of temperature non-linearly.11,12",
        "texts": [],
        "surrounding_texts": [
            "Because of the symmetry of physical model and load, half of the physical model is selected to facilitate the analysis. The model is densely meshed in laser distribution region, and the grid dimension is 3?560?856 0?3 mm. The substrate is meshed using the element of Solid70 in ANSYS, while the element of SURF152 is also used on the substrate\u2019s surface under the consideration of thermal convection and thermal radiation of heated surface. Figure 1 shows the element model after meshing. Light intensity distribution of the laser is Gaussian distribution; the laser average power is used as surface hot resource, which can be expressed as {k LT Lz ~g P pR2 r\u0192R (5) where g is the effective utilisation rate of laser power, P is the laser average power, R is the beam radius and r is the distance to the centre of beam. The model is built under the assumption that the material is isotropic. The flow of fluid in the molten pool and gasification of material are omitted. ANSYS Parametric Design Language (APDL) is used to assist finite element calculation. Latent heat of fusion is obtained by defining the enthalpy under different temperatures. Loading different laser waves is achieved by defining functions in ANSYS software. Numerical simulation results and analysis Influence of various waveforms on temperature and temperature gradient in molten pool As for three kinds of waveforms, when they have the same average output power, the influence of waveforms 1 Finite element modelling of direct metal forming a continuous wave; b sine wave; c square wave 2 Different waveforms a curve of temperature variation in the molten pool; b curve of temperature gradient variation in the molten pool 3 Dynamic curves of three kinds of waveforms in molten pool 72 Materials Technology 2011 VOL 26 NO 2 D ow nl oa de d by [ U ni ve rs ity o f L et hb ri dg e] a t 1 7: 04 0 2 A pr il 20 16 on temperature and temperature gradient in the molten pool was studied. The parameters are listed in Table 1. From the simulation calculation, the dynamic processes of temperature and temperature gradient in the molten pool by three kinds of waveforms are shown in Fig. 3. From Fig. 3a, the temperature in the molten pool increases quickly and then gradually stabilises when heated by CW laser. However, the temperature in the molten pool fluctuates periodically when heated by sine and square waveforms. From Fig. 3b, the temperature gradient in the molten pool is basically unchanged when heated by CW laser. However, when heated by sine and square waveforms, the temperature gradient in the molten pool fluctuates periodically. As to CW laser, the energy density is fixed, and the heating process parameters such as scanning velocity are fixed, so the temperature and temperature gradient in the molten pool changed a little. For sine and square waveforms, the laser power changed with time, especially square wave power changed greatly, which leads to the fluctuations of the temperature in the molten pool and great temperature gradient in the molten pool. By comparing the temperature gradient in the molten pool heated by above three kinds of waveforms, molten pool had the greatest temperature gradient when heated by square waveform, which had fast heating and cooling speeds. 4 Schematic of temperature distribution in molten pool and path A\u2013B 5 Influence of various waveforms on vertical temperature distribution in molten pool 6 Comparison of temperature field distribution of molten pool heated by a continuous wave and b square wave 7 Comparison of vertical temperature distribution in mol- ten pool heated by continuous wave and square wave D ow nl oa de d by [ U ni ve rs ity o f L et hb ri dg e] a t 1 7: 04 0 2 A pr il 20 16 Influence of various waveforms on temperature distribution in molten pool When the average output power was fixed, the influence of various waveforms on temperature distribution in the molten pool was studied. The parameters are listed in Table 1. When heated by CW laser, the temperature distribution of molten pool and the path A\u2013B are shown in Fig. 4. Through extracting the temperature data from the path A\u2013B, the vertical temperature distribution of molten pool heated by various kinds of waveforms is shown in Fig. 5. Thermal efficiency heated by square waveform is greater than that heated by CW from Fig. 5; however, when they have the same average output power, it has greater temperature gradient when heated by square waveform compared with the CW. Because of the greater thermal efficiency of the square waveform, when they have the same average output power, it can make a simulation by decreasing the average output power of square waveform and setting the appropriate parameters, which are listed in Table 2. The comparison of the temperature field distribution in the molten pool heated by CW and square wave laser at the same time is shown in Fig. 6. The right part of Fig. 6 is vertical chart, and the left is M\u2013M crosssection. From the cross-section, the molten pool heated by square wave has smaller heat affected zone compared with CW laser, which is good for the improvement of forming precision. Through extracting the temperature data from the path A\u2013B, the vertical temperature distribution in the molten pool heated by CW and square wave is shown in Fig. 7. From Fig. 7, we know that the molten pool heated by square wave with smaller average output power can achieve the same efficiency as that heated by CW, and the temperature decreased rapidly in the longitudinal direction of molten pool when heated by square wave, so it has smaller heat affected zone. The laser thermal efficiency of square wave depends on peak power rather than average power, which provides a theoretical basis for the improvement of forming precision. 8 Schematic of direct metal forming system 9 Comparison of experimental and simulation results Table 3 Process parameters and thickness of thin walled parts No. Waveform Average power P, W Repetition frequency f, Hz Duty ratio t, % Thickness, mm 1 Continuous wave 200 \u2026 \u2026 0.60 2 Sine wave 150 300 500 0.54 3 Square wave 150 300 500 0.48 10 Different cladding sections formed by continuous wave and square wave 11 Thin walled parts formed by a continuous wave and b square wave 74 Materials Technology 2011 VOL 26 NO 2 D ow nl oa de d by [ U ni ve rs ity o f L et hb ri dg e] a t 1 7: 04 0 2 A pr il 20 16"
        ]
    },
    {
        "image_filename": "designv11_100_0003670_amr.479-481.670-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003670_amr.479-481.670-Figure3-1.png",
        "caption": "Figure 3 shows the mechanical analysis of wheel hub bearing when vehicle is under braking with turning.",
        "texts": [],
        "surrounding_texts": [
            "Under the Condition of Driving with Turning. When vehicle is under driving with turning, the wheel hub bearings mainly receive the force in two directions, axial and radial. Through the analysis of load source it is clear that, the axial force mainly comes from lateral force on wheel yF , while the radial force is from three aspects, driving force on wheel DF , vertical force from ground ZF and also yF . In order to make the question more clear and simple, before we analyze the load condition of the bearing, firstly we propose the assumptions as follows: tire-road friction is good, longitudinal and lateral force don\u2019t reach the limit, don\u2019t consider the dynamic change of vertical load from road roughness, ignore effects from suspension pin, lateral force don\u2019t drift the center line of wheel, four wheels are all wheel hub motors, all motors\u2019 output torque are equal. According to the state of vehicle driving with turning, for both front and rear wheels of vehicle, ignore the load variation on both left and right wheels, and the lateral force from unsprung mass, but consider the location of gravity center moving to back, we get equations of the load on wheel [4]: 2 sin / 4 21.15 d D C A udu F m mg dt \u03b4 \u03b1  \u22c5 \u22c5 = + \u00b1    (1) ( ) ( )2 2 gr Z f r f r m du dt hmgl F l l l l = \u2212 + + (2) ( )2 s r y f r m l dv F dtl l = \u22c5 + (3) Where \u03b4 is conversion coefficient of vehicle mass, m is vehicle mass, sm is sprung mass, dC is coefficient of air resistance, u is longitudinal velocity, A is Car windward acreage, \u03b1 is road slope angle, fl is distance between gravity center and front axle, rl is distance between gravity center and rear axle, gh is height of gravity center, v is lateral velocity. external bearings. Where iF is radial force of internal bearing, oF is radial force of external bearing, e is offset distance between external bearing and wheel, l is center distance between internal and external bearings, R is wheel radius. 2 2 i D Z y e e R F F F F l l l    = + +        (4) 2 2 o D Z y l e l e R F F F F l l l \u2212 \u2212   = + \u2212        (5) Axial force of wheel hub bearing mainly comes from lateral force on wheel, which splits into two parts, axial force ' yF , in the direction of bearing axle and moment. This ' yF is equal to lateral force yF , assuming that the axial force of internal and external bearings iyF , oyF are equal. ( ) ' 4 1 2 iy oy r f r y ml dv dtl F F l F \u22c5 + = = = (6) Under the Condition of Braking with Turning. When vehicle is under braking with turning, the condition is very similar. Wheel hub bearings suffer the force in two directions, axial and radial. Axial force comes from lateral force, while radial force comes from four aspects, braking force B F , vertical force Z F , tangential force from brake caliper F\u00b5 and lateral force yF . 2 sin / 4 21.15 d B C A udu F m mg dt \u03b4 \u03b1  \u22c5 \u22c5 = \u2212    \u2213 (7) ( ) ( )2 2 gr Z f r f r m du dt hmgl F l l l l = + + + (8) ( )2 s r y f r m l dv F dtl l = \u22c5 + (9) At the same time, according to the principle of moment balance, we get the format of tangential force from brake caliper F\u00b5 , where r is effective radius of brake disc. 2 sin 21.15 dB C A uF R du R F m mg r dt r \u00b5 \u03b4 \u03b1  \u22c5 \u22c5\u22c5 = = \u2212 \u22c5    \u2213 (10) According to the condition of driving with turning we calculate the radial force of both bearings as follows, while axial force is just the same as the equation under driving with turning, where 1 l is offset distance between internal bearing and brake disc, \u03b2 is angle between brake caliper and Z axis. ( ) ( ) 2 2 1 1cos sin yB BB Z i RFl l RF l l RFeF eF F l l r l l r l \u03b2 \u03b2+ +    = + + \u2212 +   \u22c5 \u22c5    (11) 22 1 1cos sin yB B o B Z RFl RF l RFl e l e F F F l l r l l r l \u03b2 \u03b2 \u2212 \u2212 = \u2212 + + \u2212  \u22c5 \u22c5    (12)"
        ]
    },
    {
        "image_filename": "designv11_100_0003161_amm.105-107.244-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003161_amm.105-107.244-Figure4-1.png",
        "caption": "Figure. 4 Visualization of first mode shape",
        "texts": [
            " The natural frequencies and corresponding mode shapes were calculated \u2013 see Table 1, column \u201cCoupling same\u201d. By realization of steps i) \u2013 iii) both real and unreal modal shapes and natural frequencies for all possible configurations of connections (couplings) in longitudinal direction were calculated. The results are summarized in Table 1. Only such results can be considered as real where the tension force in longitudinal direction between roller and guiding U-profile is not transferred. This fact was found out on the base of visualization of particular mode shapes. The example is depicted in Fig. 4 which presents the first mode shape for the coupling configuration \u201cCoupling everywhere\u201d. The mutual position of tables and pillars requires the presence of tension forces, hence the modal shape is considered as unreal. The first real mode shape was then evaluated from the column \u201cCoupling opposite\u201d with frequency 6.5Hz. The mode shapes and natural frequencies for all described configurations of boundary conditions could be divided into two groups \u2013 real and unreal. The calculated real values of mode shapes and natural frequencies were consequently sorted in ascending order (marked in Table 1 by number of natural frequency in brackets) and could be used for the description of dynamical behavior of lifting platform"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003120_icate.2012.6403418-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003120_icate.2012.6403418-Figure6-1.png",
        "caption": "Fig. 6. Comparison between the field lines for the motor of type M3AA 71 B2/4, 3000 rpm at t=0,5 s for winding : a) Double layer b) Simple layer",
        "texts": [],
        "surrounding_texts": [
            "The analysis is performed similarly for speed of 3000 rpm. In the first stage is comparing the distribution of flux lines in the motor at t = 0.5 s. The normal component of airgap magnetic flux density has the amplitude 1,2 T for motor with double layer winding and 1,4 T in the case of motor with single layer winding. Electromagnetic torque and speed variation is shown comparatively in the following figure: From the analysis of the above figure is observed that for both types of winding, the electromagnetic torque records pulsations which are found also in the motor speed to the stable function. In the case of motor with double layer winding pulsations amplitude is lower with 2 Nm than the pulsation torque amplitude of double layer winding motor. Magnetic pressure is calculated on the stator surface for each of the two types of windings. a) For double layer winding In the case of double layer motor windings of the amplitude average magnetic pressure is recorded all the time at the value of 0.94N. The most important harmonics orders are those presented in Table III. TABLE III. ORDER AND FREQUENCY OF MAJOR PEAKS FOR THE MOTOR OF TYPE M3AA 71 B2/4, 3000 RPM DOUBLE LAYER WINDING The order of harmonics Freqency [Hz] 11 550 12 600 14 700 18 900 26 1300 38 1900 39 1950 50 2500 62 3100 80 4000 81 4050 b) For single layer winding Fig. 11. The magnetic pressure exerted on the stator surface of the M3AA 71 B2/4 induction motor, single layer winding, 3000 rpm and amplitude of the first 90 harmonics In this case the amplitude average of the magnetic pressure has all the time the value of 0.94N. The most important peaks are recorded at the frequencies presented in Table IV. From the comparison of both cases results significant amplitude harmonics, common to the two types of windings, and namely the harmonics presented in Table V. III. CONCLUSIONS Modelling of the magnetic noise was performed at two speeds 1500 rpm and 3000 rpm using the finite element software FLUX 2D. It was studied the influence of magnetic pressure on each type of winding and it has been shown the significant harmonics which appear in this case presented in Table I,II,III and IV. Following this study results that for the analyzed motors can be used the simple layer winding without exceeding the noise level requested by standards. This study was performed with the purpose to reduce time and costs required to manufacture three-phase induction motor of type M3AA 71 B2 /4 through the practical implementation practice of the obtained solutions results, and as well to observe the transitory phenomenon that occurs during operation working. ACKNOWLEDGMENT We are grateful to Electroprecizia Sacele enterprise all the support given to accomplish the measurements for the present article. REFERENCES [1] Ionescu R. M., Contributions to modeling and analysis of the noise of variable speed induction machines, PhD Thesis, \u201cTransilvania\u201d University of Brasov , Romania, 2011 [2] Scutaru G., Peter I. The Noise of the electrical induction motors with squirrel-cage rotor (in Romanian language), Publishing house LUX LIBRIS, Bra\u015fov, 2004. [3] Ionescu R. M, Scutaru G. , Peter I, Analysis the magnetic noise of the induction motor with speed control, University Publishing House 'Transylvania' ,Brasov, 2011 [4] J. Gieras, C. Wang and J. Cho Lai, Noise of polyphase electric motors, Taylor&Francis, 2006. [5] J. Le Besnerais, V. Lanfranchi, G. Friedrich, M. Hecquet and P. Brochet, \u201cPrediction of audible magnetic noise radiated by adjustable speed drive induction machines\u201d, Proceedings of Electric Machines and Drives Conference, IEMDC 2009, Miami, Florida, USA, 2009, pp. 902-908. [6] Antti Laiho, \u201cElectromechanical modelling and active control of flexural rotor vibration in cage rotor electrical machines \u201d PhD Thesis,Helsinki University of Technology,Helsinki, 2009"
        ]
    },
    {
        "image_filename": "designv11_100_0003617_amr.295-297.1631-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003617_amr.295-297.1631-Figure1-1.png",
        "caption": "Fig. 1 Plot of spur gear forging die structure",
        "texts": [
            "pur gear, the reason that the tooth corner was difficult to fill is revealed. Changing the motion mode of the floating die will change the friction condition between billet and floating die, as well as the filling situation of tooth corner. Finally, the scheme was further validated by DEFORM-3D, and the results show that the validity of floating die to tooth corner filling. Precision forging spur gear is to produce gear forms, as shown in Fig.1, which is no subsequent operations to make them fit for purpose. Compared with the conventional machining methods, the cold forging of gear has many advantages, such as great reduction of raw material and energy expenses, considerable improvement the productivity effect, and marked increase in strength values of the teeth because of the intact microstructure of the forged gear. In the last few decades, there has been an increasing interest in the production of gears by net-shape cold forging technique"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003543_20120215-3-at-3016.00205-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003543_20120215-3-at-3016.00205-Figure2-1.png",
        "caption": "Fig. 2. Excentricity and coordinate systems",
        "texts": [
            " Additional masses, for example lining discs, can easily be applied by a screw to vary the mass of the excentricity. The Projection Equation, see e.g. Bremer, H. (2008), is used to derive the dynamical equations of motion for spinning Timoshenko beams, leading to PDEs. As describing velocities y\u0307 for a mass element dm, the rigid body rotation \u03b1 (t), the torsional deflection \u03d1 (x, t), the longitudinal displacement u (x, t), the elastic bending deflections v (x, t) , w (x, t) both in the rotating reference frame, see Fig. 2, the bending angles \u03b2 (x, t) , \u03b3 (x, t) and the beams curvature given by the previous variables, all on the velocity level y\u0307T = [ \u03b1\u0307 u\u0307 v\u0307 w\u0307 \u03d1\u0307 \u03b2\u0307 \u03b3\u0307 u\u0307\u2032 v\u0307\u2032 w\u0307\u2032 \u03d1\u0307\u2032 \u03b2\u0307\u2032 \u03b3\u0307\u2032 ] (1) are used. They are the result of the application of a differential operator D\u0304 to the minimal velocities of the system s\u0307T = [ \u03b1\u0307 u\u0307 v\u0307 w\u0307 \u03d1\u0307 \u03b2\u0307 \u03b3\u0307 ] , y\u0307 = D\u0304 \u25e6 s\u0307. (2) By using a Ritz approach s\u0307 (x, t) \u223c= \u03a6 (x) T y\u0307R (t) (3) :=   1 0 0 0 0 0 0 0 U 0 0 0 0 0 0 0 V 0 0 0 0 0 0 0 W 0 0 0 0 0 0 0 \u0398 0 0 0 0 0 0 0 B 0 0 0 0 0 0 0 \u0393   \u00b7   \u03b1\u0307 q\u0307u q\u0307v q\u0307w q\u0307\u03d1 q\u0307\u03b2 q\u0307\u03b3   the differential operator can be applied to the shape functions y\u0307 \u223c= D\u0304 \u25e6\u03a6 (x) T y\u0307R (t) = \u03a8T y\u0307R",
            " The nonlinear mass- and gyroscopic matrices are given by Mb = \u222b B \u03a8dM\u03a8T (6) = \u222b L 0 \u03c1   2I 0 0 0 2I\u0398T 0 0 0 AUUT 0 0 0 0 0 \u2212AwV 0 AVVT 0 0 0 0 AvW 0 0 AWWT 0 0 0 2I\u0398 0 0 0 2I\u0398\u0398T 0 0 I\u03b3B 0 0 0 0 IBBT 0 \u2212I\u03b2\u0393 0 0 0 0 0 I\u0393\u0393T   dx, and Gb = \u222b B \u03a8dG\u03a8T (7) = \u03b1\u0307 \u222b L 0 \u03c1   0 0 0 0 0 0 0 0 0 0 0 0 0 0 \u2212AvV 0 0 \u22122AVWT 0 0 0 \u2212AwW 0 2AWVT 0 0 0 0 0 0 0 0 0 0 0 I\u03b2B 0 0 0 0 0 0 I\u03b3\u0393 0 0 0 0 0 0   dx. The rotor is characterized by a cross section area A, density \u03c1, length L and second moment of inertia I. The same procedure is used to calculate the matrices for the excentricity m\u01eb. This paper deals only with the linear momentum of mass m\u01eb, therefore the angular momentum is neglected. The position vector to m\u01eb, see Fig. 2 again, written in the spinning reference frame, is given by Rrc,\u01eb = R [ x+ u v w ] x=x\u01eb\ufe38 \ufe37\ufe37 \ufe38 rc +ARE E [ 0 s 0 ] . (8) The variable s describes the distance between the beam\u2019s rotationsal axis and the mass m\u01eb and x\u01eb is the position in longitudinal direction. Finally ARE denotes the transformation matrix from the element fixed to the reference system. It can be linearized to ARE = I+ \u03d5\u0303, (9) where I is the identity matrix and \u03d5\u0303 the spin matrix of the deflection angle \u03d5 = [\u03d1 \u03b2 \u03b3] T x=x\u01eb . (10) Calculating the bodyintegral B yields the nonlinear mass matrix for the excentricity M\u01eb = m\u01eb   m\u03b1\u03b1 0 m\u03b1v m\u03b1w m\u03b1\u03d1 0 0 0 muu 0 0 0 0 mu\u03b3 mv\u03b1 0 mvv 0 0 0 mw\u03b1 0 0 mww mw\u03d1 0 0 m\u03d1\u03b1 0 0 m\u03d1w m\u03d1\u03d1 0 0 0 0 0 0 0 0 0 0 m\u03b3u 0 0 0 0 m\u03b3\u03b3   x=x\u01eb (11) m\u03b1\u03b1 = (s+ v) 2 + (s\u03d1+ w) 2 m\u03b1v =mT v\u03b1 = \u2212 (s\u03d1+ w)VT m\u03b1w =mT w\u03b1 = (s+ v)WT m\u03b1\u03d1 =mT \u03d1\u03b1 = ( s2 + sv ) \u0398T muu =UUT mu\u03b3 =mT \u03b3u = \u2212sU\u0393T mvv =VVT mww =WWT mw\u03d1 =mT \u03d1w = sW\u0398T m\u03d1\u03d1 = s2\u0398\u0398T m\u03b3\u03b3 = s2\u0393\u0393T , and the gyroscopic matrix G\u01eb = m\u01eb   g\u03b1\u03b1 0 g\u03b1v g\u03b1w g\u03b1\u03d1 0 0 0 0 0 0 0 0 0 gv\u03b1 0 0 gvw gv\u03d1 0 gw\u03b1 0 gwv 0 0 0 0 g\u03d1\u03b1 0 g\u03d1v 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0   x=x\u01eb (12) g\u03b1\u03b1 = (s+ v) v\u0307 + (s\u03d1+ w) w\u0307 + ( sw + s2\u03d1 ) \u03d1\u0307 g\u03b1v = (s+ v) \u03b1\u0307VT g\u03b1w = (s\u03d1+ w) \u03b1\u0307WT g\u03b1\u03d1 = ( sw + s2\u03d1 ) \u03b1\u0307\u0398T gv\u03b1 =\u2212V ( s\u03d1\u0307+ w\u0307 + (s+ v) \u03b1\u0307 ) gvw =\u2212gT wv = \u2212VWT \u03b1\u0307 gv\u03d1 =\u2212gT \u03d1v = \u2212sV\u0398T \u03b1\u0307 gw\u03b1 =W (v\u0307 \u2212 (s\u03d1+ w) \u03b1\u0307) g\u03d1\u03b1 =\u0398 ( sv\u0307 \u2212 ( sw + s2\u03d1 ) \u03b1\u0307 ) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001269_978-3-540-33461-3_5-Figure11.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001269_978-3-540-33461-3_5-Figure11.7-1.png",
        "caption": "Fig. 11.7. An unmarked arrow indicates that the inclusion is proper, while a question mark indicates that it is an open problem whether the corresponding inclusion is proper. For those classes that are not connected via directed paths in the diagram it is open whether any inclusions hold.",
        "texts": [
            " An ATPDA is called shrinking, if the underlying TPDA is shrinking. By sATPDA we denote the class of all shrinking ATPDAs. As each sTPDA is a sATPDA without universal states, we see that GCS \u2286 L(sATPDA) holds. On the other hand, the Gladkij language LGL := {w#wR#w | w \u2208 {a, b}\u2217 }, which is not growing context-sensitive [1, 4, 11], is accepted by some sATPDA, implying that this inclusion is proper. Actually, the following results on the language class L(sATPDA) are known. Proposition 8. [42] (a) GCS \u2282 L(sATPDA) \u2286 DCS. (b) LOG(L(sATPDA)) = PSPACE. Figure 11.7 summarises the language classes considered here and the corresponding inclusion results. The notion of monotonicity considered in Section 11.3 is based on the requirement that, within a computation of a restarting automaton, the distance between the place where a Rewrite step is performed and the right end of the tape must not increase from one cycle to the next. Here we study the seemingly symmetric notion of left-monotonicity. We say that a sequence of cycles Sq = (C1, C2, . . . , Cn) of an RLWWautomaton M is left-monotone if Dl(C1) \u2265 Dl(C2) \u2265 \u00b7 \u00b7 \u00b7 \u2265 Dl(Cn)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.2-1.png",
        "caption": "Figure 12.2 A rotating arm.",
        "texts": [
            "188) are the most common sets: B G\u03c9B = \u03c91 \u0131\u0302 + \u03c92j\u0302 + \u03c93k\u0302 = \u03d5\u0307e\u0302\u03d5 + \u03b8\u0307 e\u0302\u03b8 + \u03c8\u0307 e\u0302\u03c8 = \u23a1 \u23a3 sin \u03b8 sin \u03c8 cos \u03c8 0 sin \u03b8 cos \u03c8 \u2212 sin \u03c8 0 cos \u03b8 0 1 \u23a4 \u23a6 \u23a1 \u23a3 \u03d5\u0307 \u03b8\u0307 \u03c8\u0307 \u23a4 \u23a6 (12.84) \u23a1 \u23a3 \u03d5\u0307 \u03b8\u0307 \u03c8\u0307 \u23a4 \u23a6 = 1 sin \u03b8 \u23a1 \u23a3 sin \u03c8 cos \u03c8 0 sin \u03b8 cos \u03c8 \u2212 sin \u03b8 sin \u03c8 0 \u2212 cos \u03b8 sin \u03c8 \u2212 cos \u03b8 cos \u03c8 1 \u23a4 \u23a6 \u23a1 \u23a3 \u03c9x \u03c9y \u03c9z \u23a4 \u23a6 (12.85) 3. Euler parameters as given in Equation (8.155): \u2190\u2192\u0307 e = 1 2 \u2190\u2192e \u2190\u2192 B G\u03c9B (12.86) e\u03070 = \u2212\u03c91e1 \u2212 \u03c92e2 \u2212 \u03c9ze3 (12.87) e\u03071 = \u03c91e0 \u2212 \u03c92e3 + \u03c9ze2 (12.88) e\u03072 = \u03c92e0 + \u03c91e3 \u2212 \u03c9ze1 (12.89) e\u03073 = \u03c92e1 \u2212 \u03c91e2 + \u03c9ze0 (12.90) Euler parameters also cover the angle\u2013axis and quaternion representations. Example 719 A Rotating Arm Figure 12.2 shows a rotating arm with a body coordinate frame B. The transformation matrix between B and G and the angular velocity of the arm are given as GRB = RZ,\u03b8 = \u23a1 \u23a3 cos \u03b8 \u2212 sin \u03b8 0 sin \u03b8 cos \u03b8 0 0 0 1 \u23a4 \u23a6 (12.91) B G\u03c9B = GRT B GR\u0307B = \u03b8\u0307 k\u0303 = \u03b8\u0307 \u23a1 \u23a3 0 \u22121 0 1 0 0 0 0 0 \u23a4 \u23a6 (12.92) Assuming a principal mass moment matrix BI = \u23a1 \u23a2 \u23a3 Ix 0 0 0 Iy 0 0 0 Iz \u23a4 \u23a5 \u23a6 (12.93) we have BM = B L\u0307 + B G\u03c9B \u00d7 BL = BI B G\u03c9\u0307B + B G\u03c9B \u00d7 ( BI B G\u03c9B ) (12.94) or \u23a1 \u23a2 \u23a3 Mx My Mz \u23a4 \u23a5 \u23a6 = \u23a1 \u23a3 0 0 Iz\u03b8\u0308 \u23a4 \u23a6 (12.95) Using the transformation matrix GRB , we can determine GI : GI = GRB BI GRT B = GRB \u23a1 \u23a2 \u23a3 Ix 0 0 0 Iy 0 0 0 Iz \u23a4 \u23a5 \u23a6 GRT B = \u23a1 \u23a2 \u23a3 Ix cos2 \u03b8 + Iy sin2 \u03b8 ( Ix \u2212 Iy ) cos \u03b8 sin \u03b8 0( Ix \u2212 Iy ) cos \u03b8 sin \u03b8 Iy cos2 \u03b8 + Ix sin2 \u03b8 0 0 0 Iz \u23a4 \u23a5 \u23a6 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.44-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.44-1.png",
        "caption": "Figure 3.44 Rolling maneuver of aircraft/aerodynamic missile",
        "texts": [
            " 43 A ge ne ra la ir cr af tl at er al -d ir ec tio na lc on tr ol sy st em ,i nc lu di ng th e in ne rs ta bi lit y au gm en ta tio n sy st em (S A S) ,t he m id dl e au to pi lo tl oo p, an d th e ou te r, no nl in ea r ga in sc he du le r fo r th e SA S co nt ro lle r ga in s ba se d up on fli gh tp ar am et er s The optimal control theory can be applied to terminal roll control of aircraft and missiles with aerodynamic control surfaces. Such an application involves isolating the pure rolling mode of the lateral-directional dynamics and posing the optimal control problem as a single-axis rotation with a fixed terminal time. The governing equation of motion for an aircraft in pure rolling mode with bank angle, \u03c3(t), roll rate, p(t), and aileron deflection, \u03b4A(t),9 (Figure 3.44) can be expressed as Jxxp\u0307 = Lpp + LA\u03b4A, \u03c3\u0307 = p, (3.187) which can be written in the following time-invariant state-space form with state vector, x(t) = [\u03c3(t), p(t)]T , and control input, u(t) = \u03b4A(t): x\u0307 = f(x, u) = Ax + Bu, x(0) = 0, (3.188) where A = \u2202f \u2202x = ( 0 1 0 Lp Jxx ) 9 \u03b4A(t) can alternatively be regarded as fin deflection angle for an aerodynamically controlled missile, or the exhaust guide vane deflection of a rocket creating a rolling moment. and B = \u2202f \u2202u = ( 0 LA Jxx ) u. In order to simplify notation we define a = Lp/Jxx, b = LA/Jxx"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003686_neurel.2012.6419964-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003686_neurel.2012.6419964-Figure4-1.png",
        "caption": "Fig 4. Kinematic model of manipulator ROBED03",
        "texts": [
            " The ANNprovidesfast model of ROBED03 enavit-Hartemberg (DH). These parameters rely exclusively on the geometric characteristics each. The parameters characteristics are [13]: i Rotation around 1iZ axis; id Translation along 1iZ axis; ia Translation along iX axis; i Rotation around iX axis. DH parameters of ROBED03 are given in table I TABLE I DENAVIT-HARTEMBERG PARAMETERS q d [m] a [m] q 1 0.27 0 pi/2 q 2 0 0 0.2196 0 q 3 0 0.064 0.23 0 q 4 0 -0.064 0 -pi/2 q 5 0 0.1724 0 0 The MatLab model of the robot ROBED03 described parameters is given in the Fig. 4. Limitations in GCs\u2019 values are presented in table II. - effector. results are adjustment. The ANN gives the output that set the robot near the desired point in the . Generalized coordinates obtained by ANN are provide faster and more precise achievement of desired point. The area around ANN output depends on the ANN error and represents operates the GA is limited on the area around achieved and desired point. The ANN inputs are position and orientation of end effector thatare represented by 6 parameters"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002941_kem.462-463.366-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002941_kem.462-463.366-Figure8-1.png",
        "caption": "Fig. 8 Stress distribution",
        "texts": [
            " (ii) To simulate the existence of the shaft and its abutment constraint, the nodes at the inner diameter of the gear were constrained with 3 DOFs (translations) as shown in Fig.7a. (iii) To locate the model in 3D space and to make the analysis possible, one node was fixed in all DOFs including rotational ones. The maximum principal stress due to misalignment was found to be 829 MPa and the maximum von Mises stress 927 MPa. Both the maximum stress occurred at the right corner of the keyway as shown in Fig. 8. The maximum combined stress from both components (i.e. the stress from normal operation and that from misalignment) was 856 MPa. It can be seen that the stress due to misalignment is very much higher than the stress resulting from the load in normal operating condition (829 vs 27 MPa). The maximum combined stress of 856 MPa, though not exceeding the strength of gear material of 944 MPa, is exceedingly high for design of critical mechanical components such as gears. For quenched and tempered steel, the fatigue limit is approximately one half of tensile strength, and in this case would be around 472 MPa"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.70-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.70-1.png",
        "caption": "Fig. 2.70 4PaCPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\C\\Pa",
        "texts": [
            " 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No. 58 T ab le 2. 4 L im b to po lo gy an d co nn ec ti ng co nd it io ns of th e fu ll y- pa ra ll el so lu ti on s w it h no id le m ob il it ie s pr es en te d in F ig s. 2. 73 , 2. 74 , 2. 75 , 2. 76 , 2. 77 , 2. 78 ,2 .7 9, 2. 80 ,2 .8 1, 2. 82 ,2 .8 3, 2. 84 ,2 .8 5, 2. 86 ,2 .8 7, 2. 88 ,2 .8 9, 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001944_9781118354162.ch13-Figure13.24-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001944_9781118354162.ch13-Figure13.24-1.png",
        "caption": "Figure 13.24 Effect of electrode coverage on the diffusion conditions. Cyclic voltammograms of the ferricyanide ion (5mM) in 0.1M KNO3. The electrode surface was covered by self-assembly of a thione carotenoid derivative. (1) free surface; (2) partial coverage; (3) high coverage degree. Adapted with permission from [27]. Copyright 2002 Elsevier.",
        "texts": [
            " Thus, for a short run time, each electrode develops an individual diffusion layer and the total current is a multiple of the current produced by each electrode. However, if the run time is longer (and electrode spacing is sufficiently low), individual diffusion layers will overlap and the array behaves as a macroscopic electrode, the current being proportional to the overall area of the array. An array of ultramicroelectrodes can be obtained by partially coating the surface of a normal electrode with an insulating layer. Figure 13.24 displays the behavior of such an electrode in the cyclic voltammetry of the ferricyanide ion. The uncoated electrode yields the curve 1 that is typical of linear diffusion conditions. Coverage by a selfassembled monolayer of a long, hydrophobic molecule leads to the formation of a network of pinholes where the reactant is free to access the metal surface, the pinholes being incorporated within an insulating layer. As indicated by curve 3, which corresponds to a high degree of coverage, this electrode behaves as an array of ultramicroelectrodes, the cyclic voltammograms being shaped as the curve in Figure 13"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001340_b978-081551497-8.50005-6-Figure3.47-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001340_b978-081551497-8.50005-6-Figure3.47-1.png",
        "caption": "Figure 3.47 2-D and 3-D plots of the calculated, stress-induced conductivity deviations underneath a bond pad during ball bonding. Each figure represents a different component of the conductivity tensor. Adapted from Osario et al.[77]",
        "texts": [
            " The anisotropic resistivity is a superposition of an intrinsic value and the strain-induced resistivity. 4. Parameterize the sensor geometry. 5. Create a simulation driver that can seek optimal sensor lay- outs subject to CMOS technology layout constraints. Optimality is measured as specified in Section 3.6 under the discussion of optimization. 6. Perform the simulations. By taking components of the conductivity tensor along the coordinate axes, being careful to relate these correctly to the crystal axes, we obtained the plots shown in Fig. 3.47. Equipped with this information, we were able to use the FDM solver to compute the potential field in the sensor (see Fig. 3.48) and extract the sensitivity. The simulation procedure 176 MEMS: DESIGN, ANALYSIS, AND APPLICATIONS CH03 9/9/05 8:50 AM Page 176 offered us the capability of extracting optimal values for geometry parameters. From the simulation series in which parameters W, H, and L were varied, it was demonstrated that reasonable criteria for the optimum value of these design parameters can be extracted in terms of maximum sensor sensitivity S"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.9-1.png",
        "caption": "Fig. 7.9 Axes assignment of the three-joint planar robot",
        "texts": [
            " The reason is quite clear that the last link for the robot is always the busiest body in motion, in contrast to the first link that is movable only by joint 1. This fact once again reinforces the concept that kinematics determines the structure of dynamics. To verify the second approach, let us look back at the same example as illustrated for the first approach to finding the inertial matrix W . This time we apply equation (7.30), which obviously requires the D-H convention to explicitly model the robotic kinematics. With the assignment of coordinate frames, as shown in Figure 7.9, we can readily determine its D-H table in Table 7.1. Then, all the one-step homogeneous transformation matrices can be derived as follows: A1 0 = \u239b \u239c\u239d c1 0 s1 0 s1 0 \u2212c1 0 0 1 0 0 0 0 0 1 \u239e \u239f\u23a0 , A2 1 = \u239b \u239c\u239d 1 0 0 0 0 0 1 0 0 \u22121 0 d2 0 0 0 1 \u239e \u239f\u23a0 , A3 2 = \u239b \u239c\u239d c3 0 s3 0 s3 0 \u2212c3 0 0 1 0 0 0 0 0 1 \u239e \u239f\u23a0 . To prepare the necessary homogeneous transformations towards the three sub-Jacobian determinations, first, it should be clarified that the first subrobot is only link 1, the second subrobot consists of both link 1 and link 2, while the third subrobot is the entire three-joint planar robot"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002724_s1052618811020142-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002724_s1052618811020142-Figure6-1.png",
        "caption": "Fig. 6. The positioning deviations of faceplate.",
        "texts": [
            " 2 2011 SERKOV Figure 5 depicts the block diagram of the drive, which is controlled according to the rotary table\u2013face plate program. Here, there are the same elements that are used in the drive, according to the linear coor dinate, but the rotary table is used instead of the slide, and the angle of rotation sensor is used instead of the linear motion sensor. Let us examine the faceplate motion as a set of two motions: ideal rotations around the immovable axis, and superposed small deviations \u03b4x(\u03d5rated), \u03b4y(\u03d5rated), \u03b4z(\u03d5rated) and \u03b1x(\u03d5rated), \u03b1y(\u03d5rated), \u03b1z(\u03d5rated). Figure 6 depicts the deviations in the position of the faceplate 1 as the small deviations of an absolutely solid body in space (4\u2014 linear deviations \u03b4x(\u03d5rated), \u03b4y(\u03d5rated), and \u03b4z(\u03d5rated) along X, Y, and Z, and angular deviations (\u03b1x(\u03d5rated), \u03b1y(\u03d5rated), \u03b1z(\u03d5rated)) (rotations) around X, Y, and Z axes, which correspond to ISO 230 7. Here O0X0Y0Z0 is the immovable Cartesian coordinate system related to the faceplate body, O1X1Y1Z1 is the coordinate system related to the faceplate, EXC is the faceplate radial pulsation towards the X axis (line 5), EYC is the faceplate pulsation towards Y axis, EZC is the faceplate axial pulsation towards the Z axis (line 2), EAC is the rotation around the axis that is parallel to the X axis, EBC is the rotation around the axis that is parallel to Y axis, ECC is the positioning deviation according to faceplate angle of rotation, and P is the faceplate pole"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002724_s1052618811020142-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002724_s1052618811020142-Figure1-1.png",
        "caption": "Fig. 1.",
        "texts": [
            " The primary deviations of link parameters of sliding and revolute pairs, which are widely used in mod ern machine structures, for example, in a five coordinate machine of grade MC 300 [9], are examined in the present paper. The primary deviations are examined in terms of the machine\u2019s standard of accuracy, in particular, GOST (State Standard) 30027 93 (Industrial Flexible Modules and Multipurpose Drilling and Milling Machines. Standard of Accuracy). Primary Deviations of Link Parameters of Kinematic Pairs Sliding pair. Figure 1a depicts the layout diagram of a five coordinate machine of grade MC 300, which is a typical example of a machine with a step by step structure; Fig. 1b presents the structural dia gram of its load carrying system mechanisms. From Fig. 1b it can be seen that two mechanisms are used (7 is spindle). The spindle body is moved according to linear coordinates X, Y, and Z, with the help of the mechanism for moving the spindle body \u201cframe (1) \u2192 slides (2) \u2192 column (3) \u2192 spindle carrier (4)\u201d; with the help of the rotary table \u201cframe (1) \u2192 faceplate body (5) \u2192 faceplate (6)\u201d the spindle body is moved according to angular coordinates B and C. Three sliding pairs are used in the mechanism for spindle body motion, and two revolute pairs are used in the rotary table mechanism",
            " 4b presents the rotary table of globe structure for the MC 700 machine (1 is the faceplate, and 2 is the faceplate body). The faceplate is intended for rotating the processing component according to the C coordinate. The faceplate body is rotated (depending on the table structure) either around the Y axis (angular coordinate B) (Fig. 4a) or around the X axis (angular coordinate A) (Fig. 4b). The block diagram of the faceplate rotary mechanism, according to two independent degrees of freedom, is the same for both structures of the rotary table and is presented in Fig. 1b. 110 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 40 No. 2 2011 SERKOV Figure 5 depicts the block diagram of the drive, which is controlled according to the rotary table\u2013face plate program. Here, there are the same elements that are used in the drive, according to the linear coor dinate, but the rotary table is used instead of the slide, and the angle of rotation sensor is used instead of the linear motion sensor. Let us examine the faceplate motion as a set of two motions: ideal rotations around the immovable axis, and superposed small deviations \u03b4x(\u03d5rated), \u03b4y(\u03d5rated), \u03b4z(\u03d5rated) and \u03b1x(\u03d5rated), \u03b1y(\u03d5rated), \u03b1z(\u03d5rated)",
            " With respect to the machine, the load carrying system is not only the frame, body elements, unit com ponents for holding, and motion of tools and the processing component, but also the mechanisms for motion transforming, including drives (electrical, hydraulic, pneumatic, etc.). For the machine mechanisms of step by step structure [1], it is possible to separate the complicated load carrying system of the machine into separate subsystems, which move according to every coordinate. From the structural diagram (Fig. 1b) it is seen that two mechanisms are used in the machine. One mech anism is for the spindle body motion \u201cframe (1) \u2192 slide (2) \u2192 column (3) \u2192 spindle head (4)\u201d according to linear coordinates X, Y, and Z. For the examined case, the \u201cX\u201d subsystem (frame (1) \u2192 slide (2)) includes the \u201cY\u201d subsystem (slide (2) \u2192 column (3)), which includes the \u201cZ\u201d subsystem (column (3) \u2192 spindle head (4)). The second mechanism is for rotating the table \u201cframe (1) \u2192 faceplate body (5) \u2192 faceplate (6)\u201d according to two angular coordinates: C and B"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003082_icems.2013.6754391-Figure16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003082_icems.2013.6754391-Figure16-1.png",
        "caption": "Fig. 16. Distribution of magnetic flux density (6 poles\u21922 poles).",
        "texts": [],
        "surrounding_texts": [
            "Fig. 17 shows the efficiency map of 6-pole and 2-pole motors. When the maximum voltage of inverter is 165V, the maximum efficiency of a 6-pole motor is about 91% from 1000rpm to 2500rpm. The maximum efficiency of a 2-pole is about 93% from 1400rpm to 3800rpm."
        ]
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.11-1.png",
        "caption": "FIGURE 1.11. A sample of tire tread to show lugs and voids.",
        "texts": [
            " The lugs are the sections of rubber that make contact with the road and voids are the spaces that are located between the lugs. Lugs are also called slots or blocks, and voids are also called grooves. The tread pattern of block-groove con gurations a ects the tire\u2019s traction and noise level. Wide and straight grooves running circumferentially have a lower noise level and high lateral friction. More lateral grooves running from side to side increase traction and noise levels. A sample of a tire tread is shown in Figure 1.11. Tires need both circumferential and lateral grooves. The water on the road is compressed into the grooves by the vehicle\u2019s weight and is evacuated from the tireprint region, providing better traction at the tireprint contact. Without such grooves, the water would not be able to escape out to the sides of the wheel. This would cause a thin layer of water to remain between the road and the tire, which causes a loss of friction with the road surface. Therefore, the grooves in the tread provide an escape path for water"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.47-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.47-1.png",
        "caption": "Figure 12.47 Free-body diagram of a 2R planar manipulator.",
        "texts": [
            "712): \u23a1 \u23a2 \u23a3 QX QY QZ \u23a4 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a3 0 0 ( I3 + (m1 + m2) r2 x ) \u03b8\u0308 + (m1 + m2) lg cos \u03b8 \u23a4 \u23a5 \u23a6 (12.713) Substituting rx from (12.691) provides the required driving torque Q0: Q0 = QZ = [Iz + m1 ( rx \u2212 l 2 )2 + m2 (l \u2212 rx) 2 + (m1 + 2m2) 2 4 (m1 + m2) l2] + (m1 + m2) gl cos \u03b8 (12.714) Example 761 2R Planar Manipulator Newton\u2013Euler Dynamics The 2R planar manipulators are applied controlled multibodies that can be seen in many robotic designs. An example of a 2R manipulator and its free-body diagram are shown in Figure 12.47. Assume G ( I\u0302 , J\u0302 , K\u0302 ) = B0 is the global coordinate frame of the manip- ulator. The driving torques of the actuators are parallel to the Z-axis and are indicated by Q0 and Q1. The Newton\u2013Euler equations of motion for the first link are 0F0 \u2212 0F1 + m1g J\u0302 = m1 0a1 (12.715) 0Q0 \u2212 0Q1 + 0n1 \u00d7 0F0 \u2212 0m1 \u00d7 0F1 = 0I1 0\u03b11 (12.716) and the equations of motion for the second link are 0F1 + m2g J\u0302 = m2 0a2 (12.717) 0Q1 + 0n2 \u00d7 0F1 = 0I2 0\u03b12 (12.718) So, there are four equations for four unknowns F0, F1, Q0, and Q1",
            "720) x = \u23a1 \u23a2\u23a2 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2 \u23a3 F0x F0y Q0 F1x F1y Q1 \u23a4 \u23a5\u23a5 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5 \u23a6 b = \u23a1 \u23a2\u23a2 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2 \u23a3 m1a1x m1a1y \u2212 m1g 0I1\u03b11 m2a2x m2a2y \u2212 m2g 0I2\u03b12 \u23a4 \u23a5\u23a5 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5 \u23a6 (12.721) The matrix A is constant. Having the column matrix b at any time, we can calculate the column matrix x to have the joint force and torques. Example 762 Actuator Torques of a 2R Manipulator In multibody dynamics, we usually do not need to find the joint forces. Actuator commands, in this case the joint torques, are more interesting because we use them to control a multibody. In Example 761 we found four equations for the joint force system of the 2R manipulator shown in Figure 12.47: 0F0 \u2212 0F1 + m1g J\u0302 = m1 0a1 (12.722) 0Q0 \u2212 0Q1 + 0n1 \u00d7 0F0 \u2212 0m1 \u00d7 0F1 = 0I1 0\u03b11 (12.723) 0F1 + m2g J\u0302 = m2 0a2 (12.724) 0Q1 + 0n2 \u00d7 0F1 = 0I2 0\u03b12 (12.725) We may eliminate the joint forces F0 and F1 from the four equations of motion (12.722)\u2013(12.725) and reduce the number of equations to two for the two torques Q0 and Q1. Eliminating F1 between (12.724) and (12.725) provides 0Q1 = 0I2 0\u03b12 \u2212 0n2 \u00d7 ( m2 0a2 \u2212 m2g J\u0302 ) (12.726) and eliminating F0 and F1 between (12.722) and (12.723) gives 0Q0 = 0Q1 + 0I1 0\u03b11 + 0m1 \u00d7 ( m2 0a2 \u2212 m2g J\u0302 ) \u2212 0n1 \u00d7 ( m1 0a1 \u2212 m1g J\u0302 + m2 0a2 \u2212 m2g J\u0302 ) (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001269_978-3-540-33461-3_5-Figure2.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001269_978-3-540-33461-3_5-Figure2.3-1.png",
        "caption": "Fig. 2.3. The nor latch: (a) circuit, (b) network",
        "texts": [
            " Forks are junctions with one input and two or more outputs. The environment supplies binary values to the input terminals. An external input terminal, a gate output and a fork output must be connected by a wire to a gate input, a fork input, or an external output terminal, and vice versa, except that we do not permit an input x to be connected to an output z, because x and z would be disjoint from the rest of the circuit. Note that a wire connects two points only; multiple connections are done through forks. Figure 2.3(a) shows a circuit called the nor latch. Forks are shown by small black dots (circled in this figure). Input x1 is connected to an input of nor gate y1 by wire w1, etc. In analyzing a circuit, it is usually not necessary to assume that all components have delays. Several models with different delay assumptions have been used [10]. We use the gate-delay model. In this model, in the latch of Fig. 2.3(a), gate variables y1 and y2 are assumed to have nonzero delays, and wires have zero delays. The choice of delays determines the (internal) state variables of the circuit. In our example, we have decided that the state of the two gates suffices to give us a proper description of the behavior of the latch. Having selected gates as state variables, we model the circuit by a network graph, which is a directed graph [1] (digraph) (V,E), where V = X \u222a Y \u222a Z is the set of vertices, and E \u2286 V \u00d7 V is the set of directed edges defined as follows. There is an edge (v, v\u2032) if vertex v is connected by a wire, or by several wires via forks, to vertex v\u2032 in the circuit, and vertex v\u2032 depends functionally on vertex v. The network graph of the latch is shown in Fig. 2.3(b). Suppose there are incoming edges from external inputs xj1 , . . . , xjmi and gates yk1 , . . . , ykni to a gate yi, and the Boolean function of the gate is fi : {0, 1}mi+ni \u2192 {0, 1}. (2.1) The excitation function (or excitation) of gate yi is Yi = fi(xj1 , . . . , xjmi , yk1 , . . . , ykni ). (2.2) If the incoming edge of output variable zi comes from vertex v, then the output equation is zi = v. In the example of the latch, we have the excitations and output equations: Y1 = x1 + y2, Y2 = x2 + y1; z1 = y1, z2 = y2",
            " Algorithm A\u0303 may not terminate, for example, for the nor latch with initial state 11 \u00b7 00 and inputs changing to 00. If Algorithm A\u0303 does terminate, let its result be sA\u0303. Note that sA\u0303 = S(a, sA\u0303), i.e., the last state is stable. The next theorem shows that it is possible to reduce the set of state variables and still obtain the same result using simulation with the remaining 2 Topics in Asynchronous Circuit Theory 23 variables. A set F of vertices of a digraph G is a feedback-vertex set if every cycle in G contains at least one vertex from F . To illustrate variable removal, consider the nor latch of Fig. 2.3. If we choose y1 as the only feedback variable, we eliminate y2 by substituting y1 + x2 for it, obtaining Y1 = x1 + y2 = x1 + y1 + x2 = x1 \u2217 (y1 + x2). In this way we obtain a reduced network3 N\u0307 of N . We perform similar reductions in the extended network N to obtain N\u0307. The proof of the following claim is similar to an analogous theorem proved in [7] for Algorithm A: Theorem 2. Let F be a feedback-vertex set of a network N, and let N\u0307 be the reduced version of N with vertex set X \u222aF . If Algorithm A\u0303 terminates on N, the final state of N agrees with that of N\u0307 for the state variables in F "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000982_9780471740360.ebs1320-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000982_9780471740360.ebs1320-Figure3-1.png",
        "caption": "Figure 3. Local coordinate system for the hand.",
        "texts": [
            " More meaningful to the rehabilitation professional is the relative movement between two segments. In most rehabilitation applications, angles are frequently computed using a local coordinate system (LCS) approach. This approach involves using the recorded position data to form localized coordinate systems for each body segment and then computing the relative angles between the segments. For example, in order to determine wrist joint motion, a local coordinate system is formed for both the hand and the forearm (Fig. 3). Three noncollinear points are needed for forming a local set of axes. For the hand, three common points of choice are the ulnar styloid (US), radial styloid (RS), and the third metacarpalphalangeal joint (3MP) (Fig. 4). The first step is to create a unit vector between two of the points. In general, the first axis is along the longest bone (or bony parts). For the hand, the first unit vector is formed between a point halfway between the US and RS (point 1) and the 3MP (point 2). Equations 2, 3, and 4 are used to create a unit vector between points 1 and 2",
            " Gimbal lock occurs when the first and third axes of rotation coincide when the second rotation is \u00fe 901 or 901 (for any combination of x, y, and z rotations) or 01 or 1801 (for a combination when the first and third axes are the same, for example, y-z-y). Near the gimbal lock position, measurement errors will be amplified and large inaccuracies of the first and third rotations will result. A gimbal lock situation can also lead to an indeterminate solution. To avoid this situation, knowledge of how the joint behaves for a given motion is necessary. For example, a common rotation scheme for determining wrist angles is x-y-z. The second rotation about y, or axis 2 in Fig. 3, is a measure of wrist deviation, which, for a majority of tasks, would not come close to 901 because of anatomical constraints. The relationship between the hand LCS and a forearm LCS is shown in Equation 18. IH JH KH 2 664 3 775\u00bc \u00bdRz;f \u00bdRy;y \u00bdRx;g x IF JF KF 2 664 3 775: \u00f018\u00de The equation is rearranged to place all the knowns on one side (hand and forearm LCS) and the unknowns (rotation angles) on the other side of the equation (Equation 19). Equation 19 also shows the rotation matrix [R] that results after multiplying the three individual rotation matrices"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003043_icinfa.2013.6720454-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003043_icinfa.2013.6720454-Figure1-1.png",
        "caption": "Fig. 1. Model of robot\u2019s legs",
        "texts": [
            " The seven-bar linkage mechanism model was chose for the trajectory planning of biped robot which is more efficient for computing and definite in physical meaning comparing with other frequently used algorithms like inverted pendulum model and cart-table model. What\u2019s more, the seven-bar linkage mechanism model does not need dynamic analysis which is very appropriate for trajectory planning of biped robot [9-11]. All the joints of the robot\u2019s legs are rotating joints and the modeling of them [12] is shown in Fig. 1. It includes nine degree of freedom in one leg. The XYZ coordinate system is the world coordinate system and the direction of X, Y and Z axes represent front, right and top of the robot, respectively. X0Y0Z0~X8Y8Z8 correspond with the D-H coordinate systems from joint 1 to joint 9 respectively. The transformation matrixes of each joint coordinate system named M0~M8 can be obtained by direct kinematic transformation. Coordinates of the robot\u2019s left and right feet respectively denoted by Pl and Pr can be transformed into that in the XYZ coordinate system through the inverse kinematic equations [13] The relationship of PL with Pl and transformation matrix Mi-1(\u03b8i), PR with Pr and transformation matrix Mi-1(\u03b8i) are expressed in equation (1)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001817_9781118443293.ch7-Figure7.29-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001817_9781118443293.ch7-Figure7.29-1.png",
        "caption": "Figure 7.29 Schematic structure of a bare gate ISFET pH sensor.",
        "texts": [
            " What is required is either a custom process that uses some other method to define the S/D regions or a completely different way to make ISFETs. Probably the most common method that has been developed to make CMOS compatible integrated ISFETs is to use the metallisation layers available in the process to effectively bring the gate connection up to the surface of the integrated circuit. There the top passivation layer of silicon nitride or silicon oxy-nitride is ideal to act as a pH sensitive membrane without having to alter the process. Figure 7.29 is a schematic crosssection through a CMOS compatible ISFET, sometimes known as an extended gate field effect transistor (EGFET), similar to a device described in [4]. The original concept for this type of EGFET structure was first published in [5]. The standard method for measuring the ISFET is to bias it in the linear region of operation. VDS is set to be some small, constant value so that the transistor channel is not pinched-off. Then, the current through the device, IDS, is also controlled and held at a constant value"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.38-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.38-1.png",
        "caption": "Fig. 2.38 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4CPPa (a) and 4PaPC (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology C\\P||Pa (a) and Pa||P\\C (b)",
        "texts": [
            " 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003082_icems.2013.6754391-Figure15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003082_icems.2013.6754391-Figure15-1.png",
        "caption": "Fig. 15. Distribution of magnetic flux density (2 pole\u21926 pole).",
        "texts": [],
        "surrounding_texts": [
            "Fig. 17 shows the efficiency map of 6-pole and 2-pole motors. When the maximum voltage of inverter is 165V, the maximum efficiency of a 6-pole motor is about 91% from 1000rpm to 2500rpm. The maximum efficiency of a 2-pole is about 93% from 1400rpm to 3800rpm."
        ]
    },
    {
        "image_filename": "designv11_100_0002801_gt2013-94270-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002801_gt2013-94270-Figure2-1.png",
        "caption": "Fig. 2 Rotating elastic interference fit (a) shaft (b) hub (c) interference fit after assembly",
        "texts": [
            " The cup-shaped diaphragm coupling contain two cup-shaped cylindrical sides connected by a thin quill shaft, the cylindrical sides of the cup-shaped diaphragm coupling-shaft system fit over the motor shaft and turbine shaft with an interference fit respectively as shown in Fig. 1. During operation, the torque moment is transferred from the motor shaft to the diaphragm coupling first, and then to the turbine shaft, through the interference fits. In convenience, the outer and inner parts of the interference fits are named as hub and shaft respectively in this paper. The interference fit of the motor shaft and left cylindrical side of the diaphragm cup-shaped coupling are enlarged, shown in Fig.2.Fig. 2(a) and (b) show the initial clearance 0\u03b4 between the interference fit parts before assembly. Fig.2 (c) shows the assembled hub and shaft. 2.2 Modeling In this paper, two issues are discussed, the function of the interference fit used to transfer torque moment and the large temperature change during operation. In the following modeling and numerical analysis, these two considerations are solved. A rotating annular disk is modeled and analyzed 2 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/01/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use firstly, then the rotating interference fit between shaft and hub is derived using their relations 2",
            " (5), the radial displacements at the inner and outer sides of the annular disk can be expressed as follows ( ) ( ) ( ) 2 2 2 2 3 2 2 2 1 1 1 1 3 1 4 e e e e r r i e e e r r u u r E E r r T E \u03ba \u03ba\u03c3 \u00b5 \u03c3 \u03ba \u03ba \u03c1 \u00b5 \u03ba \u00b5 \u03c9 \u03b1 =  += = \u2212 + \u2212 \u2212 \u2212   + + + \u2212 + \u2206  (8) ( ) ( ) ( ) 2 2 2 3 2 2 1 2 1 1 1 3 1 4 i e e i r r i e e e r r u u r E E r r T E \u03ba \u03ba\u03ba \u00b5 \u03c3 \u03c3 \u03ba \u03ba \u03ba\u03c1 \u00b5 \u03ba \u00b5 \u03c9 \u03ba\u03b1 =  += = \u2212 + + \u2212 \u2212   + + + \u2212 + \u2206  (9) The expressions of \u03c2 , \u03ba , uf , rf , f\u03b8 , ug , rg , g\u03b8 , ul , rl , l\u03b8 are listed in the Annex A. 2.2.2 Modeling of a rotating interference fit For the rotating interference fit, the initial clearance 0\u03b4 shown in Fig. 2 relates to the radial displacements of the shaft and hub through the following equation 0 hi seu u\u03b4 = \u2212 (10) The values of radial stress at the contact areas of the hub and shaft should be equal, sehihi se r rr r P\u03c3 \u03c3 == = = (11) Based on the expressions of Eqs (8) and (9), we can easily obtained the counterpart expressions of the shaft and hub, and then combining Eqs (10) and (11), we get 2 0a he i si p se he h h se sC C C C r T r T\u03c9\u03c3 \u03c3 \u03c3 \u03c9 \u03ba \u03b1 \u03b1 \u03b4+ \u2212 + + \u2206 \u2212 \u2206 = (12) Then the contact stress can be expressed as 2 0a i he h h se s he si p p p p p C C C r T r T P C C C C C \u03c9 \u03ba \u03b1 \u03b1 \u03b4\u03c3 \u03c3 \u03c9 \u2206 \u2212 \u2206 = + + + \u2212 (13) The radial displacement of the hub is 2 2 0 1 h a i h he hu hu he hu si hu hu h he h p p p h h he h h se s he hu he hu h he h h p h p C C C u r f g f f l r E C C C r T r T r f r f r T E C E C \u03c9\u03c2 \u03c3 \u03c3 \u03c1 \u03c9 \u03c2 \u03c2 \u03ba \u03b1 \u03b1\u03b4 \u03c2 \u03b1      = + + + +              \u2206 \u2212 \u2206 \u2212 + + \u2206     (14) eP iP \u03c9 ir er Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.17-1.png",
        "caption": "Fig. 2.17 4CRRR-type fully-parallel PMs with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology C\\R||R\\R (a) and C||R||R\\R (a)",
        "texts": [
            "1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig. 2.1g0) Idem No. 13 34. 4RRRC (Fig. 2.19a) R\\R||R||C (Fig. 2.1h0) Idem No. 21 35. 4RCRR (Fig. 2.19b) R\\C||R||R (Fig. 2.1l0) Idem No. 21 36. 4RRCR (Fig. 2.20a) R\\R||C||R (Fig. 2.1b0) Idem No. 21 37. 4RCRR (Fig. 2.20b) R||C\\R||R (Fig. 2.1c0) Idem No. 21 j\u00bc1 fj 5 5 23. Pp2 j\u00bc1 fj 5 5 (continued) In the fully-parallel topologies of PMs with coupled Sch\u00f6nflies motions F / G1G2-G3-G4 presented in this section, the moving platform n : nGi (i = 1, 2, 3, 4) is connected to the reference platform 1 : 1Gi : 0 by four spatial complex limbs with four or five degrees of connectivity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003535_eucap.2012.6205942-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003535_eucap.2012.6205942-Figure1-1.png",
        "caption": "Fig. 1 (a) Simplified model of the leaky-wave structure (b) Conical pattern formed by leaky-wave structure.",
        "texts": [
            " The reconfigurability of the whole LWS is ensured by internal reconfigurability of the bottom spiraphase-type reflectarray. II. REVISITING THE PRINCIPLE OF OPERATION The well-known basic equations that describe the operation of the leaky-wave antenna with periodic PRS and/or periodic HIS instead of metal screen are presented in [6]. However, in the present work the basic equations are formulated in terms of Floquet mode propagating in a periodic structure. To obtain these equations, a simplified model of LWS was used. A simplified model (Fig. 1 (a)) contains top and bottom infinite periodic reflectarrays separated by a distance h with square unit cell of period b. The periodicity of the LWS 978-1-4577-0919-7/12/$26.00 \u00a92011 IEEE 902 ensures that electromagnetic field between the reflectarrays can be expressed in terms of Floquet modes. Now assume that the top and bottom reflectarrays reflect the incident Floquet modes with reflection coefficients t and b, equal to tje and bje , respectively. This assumption implies that the top and bottom reflectarrays are ideal non-dissipative structures that reflect all Floquet modes with the phases of reflection coefficients equal to t and b , respectively",
            " These incremental phase shifts are determined by the direction of propagation of the Floquet mode defined by the elevation angle 0 and azimuthal angle 0 : 00 00 sinsin cossin kb kb y x (3) Substitution of (2) into (1) results in: 22 2 2 222 b n bb m bh l k yx bt (4) Thus, only Floquet modes with direction of propagation that satisfy (4) propagate between the reflectarrays. For a special case when a period b is less than a half of wavelength 2 , the analyzed structure supports propagating Floquet modes with n=0 and m=0. Substitution of (3) into (4) gives the following result: ...2,1,0,25.0cos 0 lhl bt (5) Therefore, the analyzed LWS with 2 b supports Floquet modes that travel only in the directions determined by the elevation angle 0 calculated with (5). Radiation exists when ideal top reflectarray is substituted with phase agile PRS (Fig. 1(b)). According to (5), the investigated LWS would demonstrate a conical radiation pattern with aperture angle 02 . The same conclusion is stated in [3,8]. Thus, it is possible to obtain reconfigurable conical pattern with two phase agile surfaces separated by a distance h. In this work spiraphase-type reflectarrays are used to control phases t and b of the reflection coefficients. III. GEOMETRY OF THE LWS The investigated reconfigurable LWS based on two spiraphase-type reflectarrays separated by a distance h is shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002599_eacm.8.00060-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002599_eacm.8.00060-Figure5-1.png",
        "caption": "Figure 5. \u2018Weather Clock\u2019. Design by [Sir] Christopher Wren submitted to the Royal Society on 9 December 1663, and to which Hooke suggested additional functions and improvements. This machine seems to have stimulated Hooke\u2019s interest in selfrecording instruments, in particular the \u2018Weather Clock\u2019 registering eight functions that he was trying to perfect up to 1678. This engraving, however, remains the only known picture of such a clock (Birch, 1756: Plate III)",
        "texts": [
            " 445, 5 December 1678). After Hooke\u2019s death, however, his horological \u2018executor\u2019, the Revd Dr William Derham FRS, described a weather clock design found among Hooke\u2019s papers that was similar to the instrument published in a drawing. Derham (1726: p. 41) tells us that in this design an 8-day pendulum clock advanced a paper-covered cylinder and a flat recording plate, while at each 15 min interval a punch device stamped out the current temperature and barometric pressure onto the slowly moving surfaces (Figure 5). I believe this to be the first proposed use of an automatic pen recorder. The imagination and technological foresight displayed in the weather clock were truly amazing, for when Hooke and Wren were born, within 3 years of each other in the 1630s, the weather was largely a mystery. By their \u2018postgraduate\u2019 years it had been discovered \u2013 within their own circle of experimentalist friends in Oxford \u2013 to be dependent on barometric pressure. By the late 1660s, not only had instruments been devised to record each of its key ingredients of pressure, temperature and humidity manually, but precision engineering mechanisms had also been designed to record each of them fully automatically and present them to the scientist as a trace mark on a piece of graduated paper"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003932_ccca.2011.6031540-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003932_ccca.2011.6031540-Figure1-1.png",
        "caption": "Figure 1. Posture definition of a two_wheeled mobile robot",
        "texts": [
            " This paper is organized as follows: Section 2 gives an overview of the problem statement involving the model of the mobile robot used in this work and the problem definition. Sections 3 and 4 describe the characteristics of both of the used approaches. The section 5 shows the extensions adapted for the motion planning algorithm to improve the previous results. Ultimately, we present the results of the simulation of the adapted algorithm. II. THE PROBLEM STATEMENT A two-wheeled mobile robot is considered and its structure is described in Fig. 1, where X-Y is the global coordinates and xm-ym is the local coordinates which is fixed to the robot with its center p as the origin. As shown in Fig.1, its body is of symmetric shape and the center of mass is at the geometric center p of the body. R is the radius of the wheel and L is the displacement from the center of robot to the center of wheel. The set (xo, yo) represents the position of the geometric center p in the world X-Y coordinates, and the angle \u03b8 indicates the orientation of the robot. The angle \u03b8 is taken counter clockwise from the X-axis to the xm-axis. These two fixed wheels are controlled independently by two motors, and the passive wheel prevents the robot from tipping over as it moves on a plane. In this paper, the motion of passive wheels is ignored in dynamics of the mobile robot. According to the schematic of the two-wheeled mobile robot described by Fig. 1, its kinetic equation can be described by (1) cos 0 sin 0 0 1 x v y \u03b8 \u03b8 \u03c9 \u03b8 \u23a1 \u23a4 \u23a1 \u23a4 \u23a1 \u23a4\u23a2 \u23a5 \u23a2 \u23a5= \u23a2 \u23a5\u23a2 \u23a5 \u23a2 \u23a5 \u23a3 \u23a6\u23a2 \u23a5 \u23a2 \u23a5\u23a3 \u23a6\u23a3 \u23a6 (1) Where x and y denote the velocity of the robot in the direction of X-axis and Y-axis, respectively. v denotes the linear velocity of the robot in the head direction of the robot (the xm-axis) and \u03b8 \u03c9= denotes the rotational angle velocity of the robot. The value of \u03b8 is positive when the robot rotate counter clockwise and the value of \u03b8 is negative if the robot rotate clockwise"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003905_kem.523-524.598-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003905_kem.523-524.598-Figure2-1.png",
        "caption": "Figure 2 (a) shows schematic image of the measurement using AFM. For the measurement, we prepare a substrate with DNA spots printed by spotting solution (75 \u03bcM T25) and a microsphere probe by using a tip-less cantilever (spring constant is 0.39 N/m) and a 10 \u03bcm-diameter polystyrene microsphere coated streptavidin. The probe is made by attaching the microsphere with light-curing resin to the end of the cantilever (Fig.2 (b)). The sphere is immobilized by the use of a micromanipulator, which allows precise location of the sphere at the apex of the cantilever. Before the measurement, we immerse the end of the cantilever in the biotinylated ssDNA (A25) solution, in order to coat A25 on the surface of the microsphere via avidin-biotin bonds. All measurements are carried out in a liquid cell filled with PBS buffer which is fixed at ambient temperature (17 2 \u00b0C). The spring constant is not calibrated. Force measurements are obtained between glass surface and the microsphere probe, and between a DNA spot and the same probe. Force curves are recorded 10 times at different positions on a substrate. After the measurements, we calculate the average of adhesion force.",
        "texts": [
            "6\u00d710 9 [microspheres/mL]. Measurement of the total binding force acting on a microsphere. If the efficiency of fabrication of microspheres is grown depending on increase of density of ssDNAs at DNA spots, there is a strong possibility that our method enables to fabricate larger micro components to substrate. In order to investigate the range of component\u2019s size that can be assembled by the proposed method, we measured the total binding force acting on a microsphere using AFM (Atomic Force Microscope). Fig.2 (a) Schematic diagram of the measurement using AFM. (b) Photograph of a 10 \u03bcm-diameter polyethylene microsphere attached to the end of the tip-less cantilever. Fig.3 Pictures of 1 \u03bcm-diameter polystyrene microspheres coated A25 assembled at DNA spot (T25). (a) A few minutes later. (b) 3 hours later. (c) 6 hours later. Fig.4 Relationship concentration of DNA in spotting solution and number of assembled 1 \u03bcm-diameter microspheres. Figure 5 shows the result of the measurement using AFM. First, for comparison, force-distance curves between a 10 \u03bcm-diameter microsphere functionalized A25 and glass surface were recorded (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure39-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure39-1.png",
        "caption": "Fig. 39 Gear Tip chamfering with Chamfer Tool",
        "texts": [],
        "surrounding_texts": [
            "direction. Such tools come in several base geometries, and tool manufacturers such as Sandvik, P. Horn, Iscar, Ingersoll Rand make disk bodies to different diameters on which blade inserts are screwed. The CoSIMT could also be a grinding disk. These tools can be used to chamfer pretty much any gear type, given the blade angles are chosen correctly. For example, the tooth tips of a spur gear can be chamfered using a CoSIMT with 15\u02da blade angles (left, Fig. 40) or 0\u02da (right, Fig. 40) and both will give the same results. The same comment applies to helical gears (left, Fig. 41), where the cutting edges are angled relative to the CoSIMT radius, and straight bevel gears (right, Fig. 41) where the cutting edges are parallel to the CoSIMT radius."
        ]
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.10-1.png",
        "caption": "FIGURE 1.10. Ground-sticking behavior of radial and non-radial tires in the presence of a lateral force.",
        "texts": [
            " One ply is set on a bias in one direction as succeeding plies are set alternately in opposing directions as they cross each other. The ends of the plies are wrapped around the bead wires, anchoring them to the rim of the wheel. Figure 1.9 shows the interior structure and the carcass arrangement of a non-radial tire. The most important di erence in the dynamics of radial and non-radial tires is their di erent ground sticking behavior when a lateral force is applied on the wheel. This behavior is shown in Figure 1.10. The radial tire, shown in Figure 1.10( ), exes mostly in the sidewall and keeps the tread at on the road. The bias-ply tire, shown in Figure 1.10( ) has less contact with the road as both tread and sidewalls distort under a lateral load. The radial arrangement of carcass in a radial tire allows the tread and sidewall act independently. The sidewall exes more easily under the weight of the vehicle. So, more vertical de ection is achieved with radial tires. As the sidewall exes under the load, the belts hold the tread rmly and evenly on the ground which reduces tread scrub. In a cornering maneuver, the independent action of the tread and sidewalls keeps the tread at on the road"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-Figure13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-Figure13-1.png",
        "caption": "Figure 13. Arrangements of Three Planet Gears",
        "texts": [
            " [1] and [2] and how to achieve an optimized crowning was discussed in Ref. [3]. When the planet gears are equally spaced around the ring gear, the gear meshing forces will be balanced so that there will be no eccentric forces on both the ring and sun gears. The condition for equally spaced planet gears is: Integer N NN SR = + )( (1) Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 4 Copyright \u00a9 2011 by ASME m=3 m=2 m=1 The left side of Figure 13 satisfies Equation (1), but the one on the right does not. There are occasions that the gear ratio requirement will force designers to use an unbalanced arrangement. Figure 14. Arrangements of Four Planet Gears For some low speed applications, the unbalanced design can be used when a small difference in force balance is not critical. Unequal spacing does not necessarily lead to force imbalance. The four-planet design on the right side of Figure 14 is not equally spaced, but is symmetric about two axes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002378_s11232-013-0053-x-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002378_s11232-013-0053-x-Figure2-1.png",
        "caption": "Fig. 2. The composite particle configuration space: when the trajectory reaches the lower boundary, the internal particle collides with the right-hand envelope wall, and when the trajectory reaches the upper boundary, the internal particle collides with the left-hand envelope wall.",
        "texts": [
            " We then pass to the new coordinates x\u03031 = \u221a m1x1 and x\u03032 = \u221a m2x2, which we call the billiard coordinates in what follows. In the billiard coordinates, the equations of motion become \u221a m1 \u00a8\u0303x1 = 0, \u221a m2 \u00a8\u0303x2 = e2E. (2) We now consider the configuration space of the composite particle. Clearly, all possible displacements of this system are described by the inequality \u2223 \u2223 \u2223 \u2223 x\u03031\u221a m1 \u2212 x\u03032\u221a m2 \u2223 \u2223 \u2223 \u2223 \u2264 L 2 , which indicates that the internal particle cannot leave the envelope. The configuration space of states of this system is the strip with the width L\u2217 = L \u221a m2 depicted in Fig. 2. This strip is slanted with respect to the axis x\u03031 by the angle \u03b1 determined by the ratio of the envelope and internal particle masses, tan \u03b1 = \u221a m2/m1. The envelope and particle motion in accordance with Eqs. (2) thus determines the parametric curve (x\u03031(t), x\u03032(t)) in the configuration space until the instant when the particle reaches the boundary of this space (see Fig. 2). Their velocities then change in accordance with the collision laws for the envelope and internal particle masses. Using the new velocities as the initial conditions when solving Eqs. (2), we obtain a new parametrically described curve (see Fig. 2). In the billiard coordinates, reflection of the trajectory by walls of the effective billiard in the strip (or from the configuration space boundary) is absolutely elastic: the angle of reflection of the trajectory from the billiard wall is equal to the angle of incidence. This can be easily explained in physical terms. The laws of mass collisions are local and independent of the presence or absence of a field. The proof of the equality between the incidence and reflection angles in the billiard coordinates therefore repeats the proof in the absence of the electric field [2]",
            " We discretize by defining or fixing velocities immediately after trajectory collisions with the billiard walls. Between collisions, the velocity changes are universal: the envelope velocity is preserved, while the internal charged-particle velocity changes with a constant acceleration. In such a discrete time pattern, we can describe the laws of velocity changes using reflections. Let the billiard trajectory at a certain evolution step reflect at the point (x\u030310, x\u030320) from the \u201clower\u201d billiard wall (see Fig. 2). This means that the initial coordinates of the reflection point satisfy the condition x\u030320\u221a m2 = x\u030310\u221a m1 \u2212 L 2 . We let V\u03032,n and V\u03031,n denote the respective internal particle and envelope velocities immediately after the reflection. The effective particle trajectory then reaches the upper wall of the billiard. We can find the trajectory at this stage using exact solutions of Eqs. (2). Indeed, integrating rather simple Eqs. (2), we obtain the velocity dependences V\u03032(t) = e2E\u221a m2 t + V\u03032,n, V\u03031(t) = V\u03031,n, (3) and the time dependences of the coordinates x\u03032(t) = e2E 2 \u221a m2 t2 + V\u03032,nt + x\u030320, x\u03031(t) = V\u03031,nt + x\u030310",
            " Let the charge of the internal particle be e2 and of the envelope be e1. In this case, it is important to know not only the value of the charge but also its distribution over the envelope. We first assume that the charge is distributed uniformly over the envelope, which implies the absence of the Coulomb interaction between the internal particle and the envelope. As before, m2 and m1 denote the internal particle and envelope masses. The configuration space in the billiard coordinates is the strip shown in Fig. 2. Its axis direction is given by the unit vector \u03c4 tangent to the boundary. In the billiard coordinates, this vector is \u03c4 = 1\u221a m1 + m2 ( \u221a m1, \u221a m2). We also introduce the vector n normal to the configuration space boundary (see Fig. 2), n = 1\u221a m1 + m2 (\u2212\u221a m2, \u221a m1), and choose its direction from the lower boundary towards the strip. The equations of motion of the composite particle components under the external field action before the collision are \u221a m1 \u00a8\u0303x1 = e1E, \u221a m2 \u00a8\u0303x2 = e2E. (16) These equations can be interpreted as the equations of motion of the effective particle with the constant acceleration in the strip a = ( e1\u221a m1 E, e2\u221a m2 E ) . This acceleration is produced by the effective electric field a acting on a unit-mass, unit-charge particle",
            " The electric field of this potential jumps in the passage from one strip to another (see Fig. 6). We thus pass to free motion in a definite field. We mention that such a generalization of the Schwarz principle is very useful when solving a broad range of physical problems related to billiards in external fields. To solve this problem, we pass to the new coordinate system in which the billiard boundaries are parallel to one of the coordinate axes. Such a transition is not principal, but it simplifies subsequent calculations. We recall that the slant angle \u03b1 of the strip (see Fig. 2) is determined by the mass ratio, tan \u03b1 = \u221a m2 m1 , cos\u03b1 = \u221a m1 m1 + m2 , sin \u03b1 = \u221a m2 m1 + m2 . We pass to the new coordinate system rotated through the angle \u2212\u03b1 with respect to the initial one: x\u2016 = x\u03031 cos\u03b1 + x\u03032 sin \u03b1, x\u22a5 = \u2212x\u03031 sin \u03b1 + x\u03032 cos\u03b1. (24) In other words, we pass to the longitudinal and transverse coordinates. We write equations of motion (16) in the new coordinates (x\u2016, x\u22a5). We begin with the case where the envelope and one of the internal particles are uniformly charged. In this case, we obtain x\u0308\u2016 = e1 + e2\u221a m1 + m2 E = e\u221a m1 + m2 E\u2016, x\u0308\u22a5 = e2m1 \u2212 e1m2 \u221a m1m2(m1 + m2) E = e\u221a m1 + m2 E\u22a5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003448_amr.463-464.1304-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003448_amr.463-464.1304-Figure2-1.png",
        "caption": "Fig. 2. Kinematics scheme of the manipulator and the trajectory described by the point H",
        "texts": [
            " The mechanisms mobility correspond with the number of independents parameters, necessary to assure the mechanism functionality. These mechanisms can be characterized by total degree of freedom, partial or fractioned [2, 3]. The actuation of those modules is realized with mechanical \u2013 hydraulics systems, type cylinder piston (CH). The manipulator structure The mechanism works sequentially, with a single motor element (the crank A0A). The kinematics\u2019 scheme of the plane mechanism [2, 3, 4], with two degree of mobility, is presented in figure 2. 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-20/05/15,22:12:43) The characteristics dimensions are, (in conformity with figure 2): ;380;680;100 660 mmlCDmmlCGmmeGA =\u2032===== .120;440;540;620 ;240;680;220 5543 3210 mmlFHmmlEFmmlDEmmlCF mmlBCmmlABmmlAA =\u2032======\u2032= ====== The maximum rotation angle of the crank shaft 1 is 260\u00b0. In the mechanism functioning we identify two phases [4]: 1. The element 6 (with the points G, C and D) stays fixed until, trough the crank shaft rotation, the point B reach on the vertical part of the element 6, between the points C and G; 2. All the kinematics elements of the mechanism are joining rigid, also continuing the crank shaft rotation until the end position, the mechanism become as a rigid body, which rotate upon the fixed joint A0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002714_amm.275-277.736-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002714_amm.275-277.736-Figure1-1.png",
        "caption": "Fig. 1. Overview of finite element modeling for lifter",
        "texts": [
            " Lifter finite element model Based on design drawings for 15T lifter, 3-dimensional modeling for lifter is established by PRO/E 3-dimensional mapping software, and finite element modeling is conducted for equipment by adopting HyperMesh v10.0; static strength analysis for lifter is conducted by using large general finite 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: 130.237.29.138, Kungliga Tekniska Hogskolan, Stockholm, Sweden-07/07/15,01:17:00) between parts, 3-dimensional modeling for lifter is conducted. Pease refer to Figure 1 for the 3-dimensional modeling of lifter. Then lead 3-dimensional modeling into HyperMesh to conduct finite element meshing and establish finite element modeling; all structures that contribute to overall rigidity and local strength of lifter will be taken into account. To ensure the accuracy of calculation, the finite element modeling of lifter mainly consists of hexahedron solid elements. Refer to Fig. 1 for the finite element modeling of lifter. After finishing the finite element modeling of all parts, the definition of contact pairs shall also be finished. In choosing surface to surface contact pairs, there are two methods: 1. Choose solid element to solid element; 2. Generate two shells on the surface of two contacts, and then define the contact relationship of two shell elements. As the equipment parts are large, many additional calculations are needed if solid contact chosen. To make calculation easier, the calculation will become more reasonable if defined surface shell elements contact chosen"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003464_amm.397-400.176-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003464_amm.397-400.176-Figure3-1.png",
        "caption": "Fig. 3 Finite Element Analysis of the planet carrier",
        "texts": [],
        "surrounding_texts": [
            "The most critical part of the gearbox design and manufacture was the planet carrier shown in Fig. 2. This single part interfaces with almost every other component in the system and transmits torque from the first stage of the gearbox to the second. Another important function of the planet carrier was to align the planet gears axially and radially with the driven shaft while rotating independently of this shaft. The planet carrier transmitted torque from the spur gear to the planet gears and each planet gear was cantilevered off bearing pins. It was therefore essential that the carrier assembly was not only strong enough to support the applied loads but also stiff enough to allow only minimal deflections under load. Large deflections would cause misalignment between the planet gears and the ring gear causing increased vibration levels and tooth wear. The planet carrier was therefore analysed using Finite Element Analysis to ensure the required stiffness was achieved. The planet carrier was manufactured from 6061 Aluminium Alloy. Fig. 4 shows the two main bearings which locate the planet carrier on the shaft and allow it to rotate independently. A shaft collar was used to mount the bearings on the driven shaft and locate the planet carrier axially. The collar also aligned the planetary stage with the driven shaft. Any misalignment in the collar or the main bearings would be amplified due to the diameter of the carrier and would cause a significant level of precession in the other components on the planetary stage. The main carrier bearings fit over the collar and were held in place with two circlips. The collar assembly is shown in Fig. 4. The large spur gear was bolted to the carrier using three 8mm shoulder bolts. Shoulder bolts were used to ensure accurate location of the spur gear on to the carrier as any misalignment would affect its meshing with the pinion gear on the driving shaft. Three steel pins were press fit into recesses in the carrier. One of the two planet bearings on each gear fits over this pin and inside the machined hole in the planet gear as shown in Fig. 5. The other planet bearings fits into the opposite side of the planet gear with a steel bush located between the two bearings. The entire assembly is bolted together using an 8mm shoulder bolt. The accurately ground shoulder of the bolt fits tightly into the bearing bush, bearing pin and carrier ensuring that the planet gear is kept in the correct position relative to the Carrier. A cross section of the assembly is shown in Fig. 6. A small clearance on either side of each planet bearing allows some axial movement of the planet gear and reduces the sensitivity of the bearings to brinelling. Fig. 7 shows a computer aided design (CAD) model of the final gearbox design. Fig. 7 CAD model of final Planetary Gearbox Design"
        ]
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.21-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.21-1.png",
        "caption": "Fig. 3.21",
        "texts": [
            " The problem can also be solved with the aid of the work-energy theorem. The kinetic energy of the drum at time t0 = 0 (i.e., \u03d5 = 0) is given by T0 = 1 2 \u0398A \u03c92 0 . At time ts (i.e., \u03d5 = \u03d5s) we have Ts = 0. The work of the external forces done during the time interval from 0 to ts is U = \u03d5s\u222b 0 MA d\u03d5 = \u2212 \u03d5s\u222b 0 r R d\u03d5 = \u2212 r R \u03d5s. Thus, (3.26) yields \u2212 1 2 \u0398A \u03c92 0 = \u2212 r R \u03d5s \u2192 ns = \u03d5s 2 \u03c0 = \u0398A \u03c92 0 4 \u03c0 r R = \u03c92 0 4 \u03c0 \u03ba . E3.8 Example 3.8 Determine the velocity of the block (mass m1) as a function of its displacement if the system in Fig. 3.21a is released from rest. Neglect the mass of the pulley R and assume both pulleys are ideal (frictionless). 3.2 Kinetics of the Rotation about a Fixed Axis 157 Solution The only external force (the weight W ) acting on the system is a conservative force. Since the velocity has to be determined as a function of the displacement of the block, it is of advantage to use the conservation of energy: T + V = T0 + V0 . We denote the displacement of the block from the initial position with x and the corresponding rotation of the upper pulley with \u03d5 (Fig. 3.21b). Then the kinetic and the potential energy are given by T0 = 0, V0 = 0 in the initial position and by T = 1 2 m1 x\u03072 + 1 2 \u0398A \u03d5\u03072, V = \u2212m1 gx in a displaced position. The kinetic energy results from the translation of the block and the rotation of the pulley. With the kinematic relation (see Fig. 3.21c) x\u0307 = 1 2 r \u03d5\u0307 \u2192 \u03d5\u0307 = 2 x\u0307 r and the moment of inertia \u0398A = 1 2m2r 2 (see (3.22)) we find[ 1 2 m1 x\u03072 + 1 2 ( 1 2 m2 r2 )( 4 x\u03072 r2 )] \u2212m1 gx = 0 \u2192 v = x\u0307 = \u00b1 \u221a 2 m1 m1 + 2 m2 gx . In the special case of a negligible mass of the pulley (m2 m1) we obtain the velocity v2 = \u221a 2 gx of the free fall of a point mass. 3.3 3.3 Kinetics of a Rigid Body in Plane Motion Let us consider a rigid body whose particles move in the x, y-plane or in a plane parallel to it (Fig. 3.22). Then the body is said to undergo a plane (planar) motion"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002656_wcse.2013.52-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002656_wcse.2013.52-Figure1-1.png",
        "caption": "Fig. 1. QUAV-X4 coordinate system.",
        "texts": [
            " INTRODUCTION Control of under-actuated QUAV-X4 has attracted more and more researchers in recent years. Improvement of sensor technology, application of new materials, and amelioration of battery life provide necessary hardware for researching it [3]. QUAV-X4 is strongly coupled, under-actuated, and nonlinear MIMO system (4 actuators and 6 DOF), while with ingenious structure design, strong controllability and high security, it has the latent capacity to take-off, hover, fly, and land in small areas. The structure of QUAV-X4 is shown as Fig.1. Where O-xyz represents earth fixed frame and Ob-XbYbZb means body fixed frame. QUAV-X4 has four rotors located at the four ends of a cross structure respectively, which are divided into counterclockwise group(M1,M3), and clockwise group(M2,M4). If QUAV-X4 is hovering, when augmented or abated rotating speeds of four rotors simultaneously, QUAV-X4 flies up and down vertically, and it will be rolling, pitching, or yawing when varying rotating speeds of one or more rotors, then because of coupling, it will be flying linearly in the space"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002316_amm.397-400.1045-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002316_amm.397-400.1045-Figure1-1.png",
        "caption": "Figure 1. The state of contact roughness surface[6] Figure 2. The model for two parallel board[3]",
        "texts": [
            "37, Georgia Tech Library, Atlanta, USA-11/11/14,17:20:32) According to the interaction among surface roughness,contact pressure for seal and the leakage of gas that writes by Xiao Ren[3], we use the software of matlab to solve the minimum pressure for seal underring the standard of leakage and then, using finite element software of ABAQUS to simulate the process of assembly and to obtain the optimal size of piston which is corresponding with the pressure that solved by matlab. The minimum contact pressure under the standard of leakage The relationship among surface roughness, contact pressure for seal and the standard of leakage Fetching a rectangular control unit, which contains enough rough peak. The real contact state between the two rough surfaces is shown in Figure 1, The expression equation of film thickness is [3],    = ++= .int0 21 potoucht t \u03b4 \u03b4\u03b4\u03b4\u03b4 (1) Where, t\u03b4 is a random location of film thickness, \u03b4 presents the average film-thickness, 1 2,\u03b4 \u03b4 are the average distances between the face and a random roughness surface. Flow factor is used to express the impact of roughness surface on the flow by PATIR and CHENG[6]. Besides, references [6] studied the issue of fluid lubricate about incompressible fluid traverse the 3-D roughness surface under isothermal condition"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001869_9781118530009.ch9-Figure9.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001869_9781118530009.ch9-Figure9.2-1.png",
        "caption": "FIGURE 9.2 Schematic drawing of CZE with end-column AD: (a) capillary; (b) cathodic buffer reservoir and electrochemical cell; (c) carbon fiber electrode; (d) electrode assembly; (e) micromanipulator; (RE) reference electrode.",
        "texts": [
            " Moreover, it was demonstrated that electrophoretic separation voltage had a greater effect on noise when capillaries with large, rather than small, internal diameters were used [16]. End-column detection, which does not require a decoupler, has been predominantly used in CE with AD. One of the first end column detectors, which did not require a decoupler, based on the judicious choice of separation capillary diameter, which was utilized for both CD and AD was introduced by Huang et al. [21] For CD, the WE consisted of a 50mm Pt wire housed in a fused silica capillary, while for AD, the WE consisted of 10mm carbon fibers, as shown in Figure 9.2. The need for a decoupler for AD was removed by utilizing a separation capillary of just 5mm diameter, where the electrophoretic current through the capillary was very small (1\u201315 nA) and electrical interference was minimized. Because the diameter of theWEwas larger than the diameter of the capillary itself, good oxidation efficiency was observed. Because CD did not appear to be as sensitive to electrophoretic currents, larger separation capillaries were successfully interfaced to this detector without the requirement of a decoupler"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001269_978-3-540-33461-3_5-Figure14.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001269_978-3-540-33461-3_5-Figure14.5-1.png",
        "caption": "Fig. 14.5. A PN with a P-invariant (1,2,1).",
        "texts": [
            " There are, however, various subclasses of PNs for which an extended state equation is sufficient and necessary to capture reachability of the underlying PN. More will be said about this in our subsequent discussion. Place Invariant: A place invariant of PN P = (P, T, \u03d5) is a mapping InvP : P \u2192 Z (i.e., assigning weights to places) such that \u2200\u00b5, \u00b5\u2032 and t \u2208 T , if \u00b5 t \u2212\u2192 \u00b5\u2032, then\u2211 p\u2208P InvP (p)\u00b5(p) = \u2211 p\u2208P InvP (p)\u00b5\u2032(p). In words, the firing of any transition does not change the weighted sum of tokens in the PN. Consider the PN shown in Figure 14.5. It is reasonably easy to observe that (1, 2, 1) is a P-invariant. Other P-invariants include (1, 1, 0), (2, 5, 3), (\u22122, 1, 3). Note that any linear combination of P-invariants is a P-invariant. Any solution of the equation X \u00b7 [T ] = 0 is a P-invariant, where X is a row vector. For instance, t2t1t2 t1 t2 1 0 0 1 0 + 1 -1 0 0 -1 1 0 0 0 1 -1 0 0 0 -1 1 0 0 1 -1 1 2 0 0 = 0 1 2 1 0 State Equation: 0+ M x = 352 Hsu-Chun Yen ( 1, 2, 1 ) \u239b \u239d \u22121 1 1 \u22121 \u22121 1 \u239e \u23a0 = ( 0 0 ) , as (1,2,1) is a P-invariant",
            " Using the so-called Farkas Algorithm, the minimal P-invariants (i.e., bases) of a PN can be calculated. Nevertheless, in the worst case the number of minimal P-invariants is exponential in the size of the PN, indicating that Farkas Algorithm may require exponential worst-case time. Suppose \u00b5 is a reachable marking (from the initial marking \u00b50) through a firing sequence \u03c3. Clearly, \u00b50 + [T ]#\u03c3 = \u00b5. (Here \u00b50 and \u00b5 are column vectors.) Let X be a P-invariant. Then X \u00b7 \u00b5 = X \u00b7 (\u00b50 + [T ] \u00b7 #\u03c3) = X \u00b7 \u00b50 + X \u00b7 [T ] \u00b7 #\u03c3 = X \u00b7 \u00b50. Recall that in the example in Figure 14.5, (1, 1, 0) is a P-invariant. For every reachable marking \u00b5, we have \u00b5(p1) + \u00b5(p2) = \u00b50(p1) + \u00b50(p2), meaning that the total number of tokens in p1 and p2 together remain unchanged during the course of the PN computation. Hence, if the PN starts from the initial marking (1, 0, 1), then the property of mutual exclusion for places p1 and p2 can be asserted as \u00b5(p1) + \u00b5(p2) = \u00b50(p1) + \u00b50(p2) = 1 for all reachable marking \u00b5. It is easy to see that if there exists a P-invariant X with X(p) > 0, for all p \u2208 P , then the PN is guaranteed to be structurally bounded. Hence, place invariants can be used for reasoning about structural boundedness. Transition Invariant: A transition invariant of PN P = (P, T, \u03d5) is a mapping InvT : T \u2192 N (i.e., assigning nonnegative weights to transitions) such that \u03a3t\u2208T InvT (t)(\u2206(t)) = 0. In words, firing each transition the number of times specified in the Tinvariant brings the PN back to its starting marking. Again consider the PN shown in Figure 14.5 in which (2, 2) (i.e., InvT (t1) = InvT (t2) = 2) is clearly a T-invariant, so is (n, n), for arbitrary n \u2265 0. Like P-invariants, any linear 14 Introduction to Petri Net Theory 353 combination of T-invariants is a T-invariant. It is easy to see that T-invariants correspond to the solutions of the following equation: [T ] \u00b7 XT = 0 (where X is a row vector representing a T-invariant). The existence of a T-invariant is a necessary condition for a bounded PN to be live. To see this, suppose P = ((P, T, \u03d5), \u00b50) is a live and bounded PN"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002656_wcse.2013.52-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002656_wcse.2013.52-Figure4-1.png",
        "caption": "Fig. 4. blade element frame of axes.",
        "texts": [
            " In order to calculate conveniently, two important velocity ratios, inflow ratio and advance ratio , are given by, . (5) . (6) 978-1-4799-2883-5/13 $31.00 \u00a9 2013 IEEE DOI 10.1109/WCSE.2013.52 283 The component of V0 along xsoszs plate has different effects on different positions of blades. For analyzing easily, blade element frame of axes op-xpypzp is shown as Fig.3. Define as azimuth of blade, means that blade is located in the negative direction of xs axis. Select blade element dr, whose average chord length is c, and the distance from which to os is r. The blade element is shown as Fig.4. Where , represent attack and inflow angles of blade respectively, means setting angle, the relationship among , and is, \u2248 . (7) Vp and Vt are vertical and circumferential components of V0, whose values are given by, . (8) When neglecting the effect of radial airflow, blade element develops derivative lifting force dL and derivative drag dD, whose directions are vertical and parallel to the direction of airflow, and whose values are given by, Where is air density, means lifting coefficient, , a is constant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.20-1.png",
        "caption": "Figure 11.20: Diagonal distortions of liftgate e left top to right bottom due to right rear tire input.",
        "texts": [
            " The baseline response is close to a minimum value and any further increase in door stiffness would not result in much further improvement in door chucking performance. It was discussed in reference 3 that 53% of vehicle squeak and rattle problems are associated with closure openings. It was also indicated in this reference that a strong correlation exists between diagonal distortions and overall squeak and rattle performance. Therefore, diagonal distortions are used as performance metrics in determining the effect of door seal stiffness on overall squeak and rattle performance. Ten diagonals at door and liftgate openings used in the analysis are shown in Figure 11.20. The dynamic responses of each of these diagonals were computed (below 50 Hz) using the same vehicle low frequency NVH CAE model. Again, four tire excitations as described in the previous section were used in this sensitivity study. In total, 40 frequency response functions were determined. A typical response plot is shown in Figure 11.20. The peak response value in each response plot is used as the performance metric. A review of all 40 responses showed that the following seven responses are significantly higher than the remaining ones: \u2022 right front door (top of A-pillar to bottom of B-pillar due to right rear tire input) \u2022 left front door (top of A-pillar to bottom of B-pillar due to left rear tire input) \u2022 left rear door (top of B-pillar to bottom of C-pillar due to left rear tire input) \u2022 liftgate (right top to left bottom due to left rear tire input) \u2022 liftgate (left top to right bottom due to left rear tire input) \u2022 liftgate (right top to left bottom due to right rear tire input) \u2022 liftgate (left top to right bottom due to right rear tire input)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002042_9783527643981.bphot009-Figure2.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002042_9783527643981.bphot009-Figure2.1-1.png",
        "caption": "Figure 2.1 Cross-section through the sensing platelet of a fiber-optic glucose sensor. P, Plexiglas; D, dialyzing membrane; E, enzyme solution; O, O-ring; L, light guide. The arrows indicate the diffusion processes involved (G, glucose; GL, gluconolactone). The directions of the exciting light (Exc)",
        "texts": [
            " Longwave sensing is highly desirable in view of the reduced absorbance and fluorescence of blood and serum at >600 nm. Labeling usually does not strongly affect the binding constants of the enzymes. b-D-glucose\u00fe FAD!D-glucono-1; 5-lactone\u00feFADH2 \u00f02:4\u00de FADH2 \u00feO2 ! FAD\u00feH2O2 \u00f02:5\u00de D-glucono-1; 5-lactone\u00feH2O! gluconate\u00feH\u00fe \u00f02:6\u00de Changes of the intrinsic fluorescence may involve (a) the UV fluorescence of tryptophan groups [excitation/emission maxima (lex/lem) at 295/330 nm] in the protein [5], (b) the fluorescence of FAD (Figure 2.1) (lex/lem 450/520 nm) [8], or (c) NAD\u00fe (lex/lem 340/460 nm) [9]. An interesting competitive FRET (F\u20acorster resonance energy transfer) assay was presented by D Auria et al. [10]. They used the intrinsic tryptophan fluorescence of glucose kinase as donor whereas glucose derivatized with o-nitrophenyl-b-D-glucose-co-pyranoside is the acceptor. Addition of glucose increases donor emission and decreases FRET. The analytical range of these sensors is typically from 0.5 to 20mM, and rarely up to 200mM [7]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003632_amm.448-453.3365-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003632_amm.448-453.3365-Figure2-1.png",
        "caption": "Fig. 2 Analysis on the speed triangle of single-row planetary gear mechanism",
        "texts": [
            " This equation is a three-variable linear equation, with three unknown numbers, which also reflects the single-row planetary gear mechanism with two degrees of freedom. A second relation equation must be added to keep a determined transmission relationship between any two components of the planetary row. In other words, a rotating component in the single-row planetary gear mechanism with two degrees of freedom must be constrained, so that the mechanism has only one degree of freedom to achieve power transmission [4] . The analysis method of speed triangle for single-row planetary gear mechanism In the single-row planetary gear mechanism as shown in Fig.2, if the planet carrier is used as a power input component, the linear velocity is Vc; the sun wheel is fixed; the gear ring is used as a power output component, the linear velocity is Vr. The size and direction of the linear velocity Vr of the power output component gear ring can be determined according to the instantaneous center and Vc. These three components are based on the center axis of the gear train as the public rotation center, which is connected to the input Vc terminal and is extended to intersect with the output Vr; the connected extension line is called a constant speed line, and the linear velocity Vd formed by the point of intersection is a virtual linear velocity with the public rotation center as the center of circle and with the same isometric speed as the input component"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002357_ecce.2013.6647276-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002357_ecce.2013.6647276-Figure3-1.png",
        "caption": "Fig. 3. Analysis model for 3D-FEM calculation.",
        "texts": [
            " The magneto-motive force (MMF) distribution induced by the daxis current, however, may influence the PM remagnetization and, thus, may cause significant torque ripple, which affects the positioning accuracy in the rotational axis (\u03b8z). The final goal of this research is to realize a multipleDOF bearingless servomotor with a wide range of motion. As the first step of the present study, PM magnetization and its influence on torque ripple with several pole/slot combinations of an SPM motor are investigated, using threedimensional finite element method (3D-FEM) analysis. II. 3D-FEM ANALYSIS CONDITIONS Figure 3 shows an example of an analysis model and its coordinates. The outer diameter of the rotor, radial width of the PM, and air gap length are 20 mm, 2.5 mm, and 2 mm, respectively. The yoke width, tooth width, and outer diameter of the stator are 5 mm, 4 mm, and 58 mm, respectively. For size reduction of the motor, short-pitch concentrated windings around the teeth were selected for this study. A horizontal direction of the stator corresponds to the X-axis. When the rotor\u2019s N pole is aligned with the X-axis, the mechanical rotational angle of the rotor \u03b8 is defined as \u03b8 = 0\u00b0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure7-1.png",
        "caption": "Figure 7: Components of Hohlraum manipulation stage",
        "texts": [
            " It can adjust the position error of TMP. In the assembly process, the contact force of TMP and Hohlraum can be detected by 6-axis force / Proc. of SPIE Vol. 8418 841819-3 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 04/21/2013 Terms of Use: http://spiedl.org/terms r Torque sensor. The attitude of TMP can be adjusted in the insertion process. The Hohlraum manipulation stage is used for holding and adjusting Hohlraum. The stage is comprised of Rotary / goniometry stages and Hohlraum holder (Figure 7). The radial angle can be adjusted by the rotary stage. The 2-axis goniometry stages can adjust the attitude of Hohlraum to be level. The Hohlraum holder can fix and support the Hohlraum, which can avoid Hohlraum circumvolving and damaging owing to stress. The detection system online is used for detecting the position and attitude of TMP and Hohlraum. It is comprised of optical image system and image detection software (Figure 8 and 9). There is also a Keyence laser sensor for adjusting the subassemblies horizontal attitude"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.9-1.png",
        "caption": "Figure 12.9 A small mass dm at Grdm of a rigid body with a fixed point at O is under a small force df.",
        "texts": [
            " The intersection forms a trajectory, such as shown in Figure 12.8. We conclude that for a given angular momentum there are maximum and minimum limit values for possible kinetic energy. Assuming I1 >I3 >I3 (12.165) the limits of possible kinetic energy are Kmin = L2 2I1 (12.166) Kmax = L2 2I3 (12.167) and the corresponding motions are turning about the axes I1 and I3, respectively. Example 727 Alternative Derivation of Euler Equation of Motion Consider a rigid body with a fixed point as shown in Figure 12.9. A small mass dm at Grdm is under a small force df. Let us show the moment of the small force df by dm: Gdm = Grdm \u00d7 Gdf = Grdm \u00d7 Gv\u0307dm dm (12.168) The angular momentum dl of dm is equal to Gdl = Grdm \u00d7 Gvdm dm (12.169) and according to (12.3), we have dm = Gd dt dl (12.170) Grdm \u00d7 df = Gd dt ( Grdm \u00d7 Gvdm dm ) (12.171) Integrating over the body gives \u222b B Grdm \u00d7 df = \u222b B Gd dt ( Grdm \u00d7 Gvdm dm ) = Gd dt \u222b B ( Grdm \u00d7 Gvdm dm ) (12.172) However, utilizing Grdm = GdB + GRB Brdm (12.173) where GdB is the global position vector of the central body frame, simplifies the lefthand side of the integral to \u222b B Grdm \u00d7 df = \u222b B ( GdB + GRB Brdm )\u00d7 df = \u222b B GdB \u00d7 df + \u222b B G Brdm \u00d7 df = GdB \u00d7 GF + GMC (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003429_amr.422.314-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003429_amr.422.314-Figure1-1.png",
        "caption": "Fig. 1 Structure schematic diagram of the test rig and installation place of sensors in the test rig.",
        "texts": [
            " Fault Diagnosis is associated on equipment operation state and abnormal conditions make a judgment to say, in the equipment failure, not before to the operation of equipment state to forecast and the forecast; In the equipment failure, for the of the cause of the failure parts such as degree type makes a judgment and fault diagnosis of the maintenance decision task including fault A new type accelerated rolling contact fatigue test rig [6-8 ]was constructed, where a ball sample could be tested with three contact points and in pure rolling condition. A schematic diagram of the test rig is shown in Fig. 1. The three contact points were on the circumference of the same largest cross-section, and the maximum Hertz contact stresses were equal. The specimen ball rested on two rollers. The loading on the ball was applied through a driver roller, which was connected to a lever so that the coil spring load could be magnified. The two rollers were supported by two cylinder rolls and one guide roll. The guide roll has a shallow groove on its rolling track. 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: 152.14.136.77, NCSU North Carolina State University, Raleigh, USA-25/04/15,13:50:40) Test rig main rotating members eigenfrequency calculation Rotation parts eigenfrequency. According test rig operating principle, the pure rolling is occurred between the specimen balls and driving. All contact point line speed is equal. As shown in Fig 1, the rotation frequency of each member is calculated as follow[9]. f2=(D1/D2) f1 (1) f3=(D1/D3) f1 (2) f4=(D1/D4) f1 (3) f5=(D1/D5) f1 (4) Where f1 is the rotation frequency of drive roll, f2 is the rotation frequency of specimen ball, f3 is the rotation frequency of accompany roll, f4 is the rotation frequency of cylinder roll and f5 is the rotation frequency of guide roll. Where D1 is diameter of driving roll, D2 is diameter of specimen ball, D3 is diameter of accompany roll, D4 is diameter of cylinder roll and D5 is diameter of guide roll"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002988_kem.462-463.762-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002988_kem.462-463.762-Figure1-1.png",
        "caption": "Fig. 1 (a) Outlook, (b) dimension in mm of the lateral plate of rescue robot",
        "texts": [
            " Tan and Gao [2] show the powerful application of the BEM in the analysis of biomaterial interface crack problem. Lih-jier Young and Tsai [3] show the powerful application of the BEM in rough contact mixed mode with the plastic crack tip problem and gives the accuracy up to 98.11%. Boundary element method is used to model the complex resistance to the applied field of the lateral plate of rescue robot to find the stress distribution along the interface. The outlook and dimension in mm of the plate are shown in Fig. 1. Young [4] also shows the potential application of the BEM determining the effect of crack face roughness in a realistic experimental specimen. The maximum allowable stress of the plate is defined by both the normal and the shear stress of first point along the interface reach the yield stress \u03c3y and\u03c4y, respectively, where\u03c4y=\u03c3y/ 3 . The plastic displacement of the interface is obtained by increasing the loading of the plate after the first point yield. Keep increasing the loading and we can get the plastic displacement of the interface"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003485_s1068798x12060093-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003485_s1068798x12060093-Figure3-1.png",
        "caption": "Fig. 3. Safety bearing: (1) rotor; (2) shoe; C, damping coef ficient of elastic component; c, rigidity of elastic compo nent.",
        "texts": [
            " To this end, we introduce damping of the elastic links within the model of the turbine rotor. The dissipation coefficients within the elastic elements of the turbine rotor are determined separately for the rotors in the generator and turbocompressor. The dis sipation coefficient is selected from the results of pre liminary calculations, so that the aperiodicity coeffi cient of vibration of the rotor components is 0.01 for critical frequencies within the working speed range. The loading of the safety bearing is illustrated in Fig. 3. Rotor 1 is installed in a safety bearing of shoe type with gap \u03941. The shoe 2 is spring loaded in the radial direction and may move a distance \u03942 until it makes contact with the rigid housing. Initially, the turbine rotor is assumed to turn in the electromagnetic bearings relative to its geometric cen ter at the rated speed. Then the radial electromagnetic bearings are switched off. On coasting, the loss of rotor speed is determined by the drag torque in the blade system of the compressors and the turbine and by the Coulomb friction in the interaction of the rotor and the safety bearing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003464_amm.397-400.176-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003464_amm.397-400.176-Figure8-1.png",
        "caption": "Fig. 8 Planet Carrier Assembly",
        "texts": [
            " 2. The main bearings are then pressed into the planet carrier and locked in location with circlips. The accelerometer is then bolted onto the carrier before the large spur gear is accurately located in position on the carrier with the shoulder bolts. 3. The slip ring carrier is bolted onto the flange of the planet carrier and slip ring is then slid onto the carrier using six set screws to hold it in place. The assembly of the planet carrier with the spur gear and slip ring in place is shown in Fig. 8. 3. The shaft collar can then be pressed into the main bearings using the assembly tools. The shaft collar is then pressed onto the driven shaft leaving a 1 mm spacing between the carrier and the sun gear. 4. The driven shaft is then inserted into position in the gearbox housing with two bearings at each end of the shaft. 5. Fig. 9 shows both stages of the UNSW planetary gearbox inside the gearbox housing. 6. The planet gears are then bolted into place on the planet carrier. The shoulder bolts supporting the planet gears are then tightened in sequence to ensure the spur gear and the planet carrier are aligned correctly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003938_sas.2012.6166281-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003938_sas.2012.6166281-Figure2-1.png",
        "caption": "Figure 2. (a) Block diagram of the described system (b) Prototype of the gyroscopic system.",
        "texts": [
            " Since mechanical vibrations can disturb the measure of the angular rate, an amplifying hardware filter with steep frequency response is mandatory at the sensor output to cut all frequencies related to the rotation speed of the main rotor and keep the largest possible bandwidth of the sensor response. To this purpose a Sallen - Key filter with double \u2013 pole response has been designed with two coincident poles at cutoff frequency of 50 Hz; at this frequency the rotor vibrations and engine vibrations are attenuated by 6 dB and 35 dB, respectively. The active filter output voltage is then acquired by an analog to digital converter on board on DSP. Figure 2 shows the block diagram of the described system and the realized prototype board. The chosen DSP is a Microchip DSPIC 30F3013 with 16 bit core and 12 bit integrated analog to digital converter. It can operate up to 30 MIPS, but in this application the selected speed is 8 MIPS, being the system battery operated and power consumption linearly increasing with DSP clock frequency. The DSP receives three inputs and provides one output, as also shown in Figure 2 (a). Two inputs are standard (for RC helicopters applications) PPM modulated digital signals, coming from the RC receiver and corresponding to two signals set by the pilot: Stick and Gain. Stick signal is the setpoint and represents the desired yaw rate of the tail; Gain signal allows the pilot to set the dynamic response of the system, i. e. How aggressive the control must be in aerobatic maneuvers. These two inputs are digital 50 Hz square waves; the normalized information is provided by the pulse duration: 0% corresponds to 1020 \u00b5s and 100% corresponds to 2020 \u00b5s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.2-1.png",
        "caption": "Fig. 3.2 Rod pendulum, length L",
        "texts": [
            " That is, if the pendulum were pivoted at the centre of oscillation, then the previous frictionless pivot would become the new centre of oscillation, and the pendulum period, P , would be the same. The centre of oscillation is sometimes called the centre of percussion (Rawlings 1993). This is because if a moving idealised compound pendulum is stopped at the centre of percussion there is no reaction at the frictionless pivot. As an example, consider an idealised compound pendulum consisting of a rigid uniform rod, length L, with a frictionless pivot at one end (Fig. 3.2). The moment of inertia, I , is ML2/3, where M is the mass of the rod, the radius of gyration, , is L= p 3, and the distance of the centre of mass from the frictionless pivot, h, is L/2. Hence, from Eq. 3.4, the effective length, l , of the compound pendulum is 2L/3. A double rod pendulum consists of two simple rod pendulums (Fig. 2.1a) arranged in series, as shown in Fig. 3.3 (Lamb 1923; Kibble and Berkshire 1996; Baker and Blackburn 2005; Vanko\u0301 2007). The upper rigid massless rod, length l1, is suspended at its upper end from a horizontal frictionless pivot, and has a point mass, m1, at its lower end"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003665_isie.2013.6563839-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003665_isie.2013.6563839-Figure12-1.png",
        "caption": "Fig. 12. Wheelchair for Experiment",
        "texts": [
            " Starting door handling, the assist force is generated by negative impedance property, and it contributes to assisting human force to accelerate the door. When the velocity is increased over vth, in which second conditions is not satisfied, the impedance is switched to normal one with getting relative motion property as shown in Fig.7. As a result, the wheelchair is automatically followed to the opening door motion, and easily complete the operation continuously. The proposed opening-door assistance is verified by using electric wheelchair as shown in Fig.12. Two IEEE 1394 High speed cameras with 1/3inch progressive scan CCD are mounted. Two direct drive AC servo motors are employed for differential wheel drive, and they are controlled by Linux PC controller (applied RT preemption patch) . WindowsPC for image processing is also mounted on the robot. Two PCs are connected through Ethernet link. The motion control sampling is 1ms, and estimated velocity from image processing is updated at frame rate (60Hz). Several control parameters are listed in Table"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003531_mace.2011.5988269-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003531_mace.2011.5988269-Figure3-1.png",
        "caption": "Figure 3. Finite element mesh.",
        "texts": [],
        "surrounding_texts": [
            "Keywords-laser cladding; strain-stress field; finite element method\nI.\n[1- 3] [4,5]\n[6]\n10 AISI1010 , 1Cr18Ni9 (AISI302) 0.3 1Cr18Ni9 206GPa\nI",
            "III. 4 720W 1mm/s\n40 38\n40 38 43.5 1400\n43.5 1400\n1Cr18Ni9 10\n5 7 x y z\n1400\n1400",
            "5 8\nZ X Y\n[11] 8\n1 3 2 Critical Strain rate for\ntemperature drop, CST\nCST\nCST"
        ]
    },
    {
        "image_filename": "designv11_100_0002992_scis-isis.2012.6505132-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002992_scis-isis.2012.6505132-Figure4-1.png",
        "caption": "Fig. 4. Outside appearance of simulation",
        "texts": [
            "(12), we can derive the first-order configuration prediction q\u0302(t\u2217i ) of actual manipulators\u2019 configuration as follow. In this paper, it is noticed that the predictive equation q\u0302(t\u2217i ) does not include the manipulators\u2019 dynamics. q\u0302(t\u2217i ) = (1 + i\u00b7t\u0303 h )q(t)\u2212 i\u00b7t\u0303 h q(t\u2212 h) (13) In order to compare the Multi-Preview Control with predictive control, we use a 7-link manipulator for simulations, which is produced by Mitsubishi Heavy Industries named PA10. Hand tracking trajectory and given manipulator\u2019s shape are depicted in Fig. 4, target hand trajectory is predefined. In addition, the kinematics of PA10 shown in Fig. 5 is implemented in the simulator. The solid line in Fig. 4 expresses a target trajectory set to be followed. The simulation\u2019s screen shot is shown in Fig. 6. The angle of actual manipulators\u2019 link 1 and the predictive angles q\u03021(t \u2217 1), q\u03021(t \u2217 2), q\u03021(t \u2217 3) of manipulators\u2019 link 1 are respectively indicated in Fig. 7, Fig. 12, Fig. 14 and Fig. 16. The angle of actual manipulators\u2019 link 2 and the predictive angles q\u03022(t \u2217 1), q\u03022(t \u2217 2), q\u03022(t \u2217 3) of it are respectively indicated in Fig. 8 when predictive interval time is 1.2[s]. Moreover, we use Runge Kutta method to calculate current angle of actual manipulator in simulation, the interval time h of Runge Kutta is 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003110_s1068798x11060153-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003110_s1068798x11060153-Figure1-1.png",
        "caption": "Fig. 1. Rotation of a tricycle with one controllable front wheel.",
        "texts": [
            "82 Understandably, a tricycle will be less stable than a quadricycle (a two axle automobile). In contrast to an automobile, in which the tipping axis on turning passes through the centers of the contact spots of the outer wheels, the tipping axis for a tricycle passes through the centers of the contact spots of the single wheel and the wheel external to the center of rotation (Fig. 1a) [1]. In Fig. 1, O, O ' are the centers of rotation of tricy cles with rigid and elastic wheels; R is the distance from the center of rotation to the tricycle\u2019s longitudi nal symmetry axis; B, L are the track and base of the tricycle (L = a + b, where a and b are the distances from the center of mass to the tricycle\u2019s front and rear axes); F, Fx, Fy are the centrifugal force and its longi tudinal and transverse components; Y1, Y2 are the lat eral (transverse) reactions of the supporting surface; \u03b8 is the angle of rotation of the controlled wheel; \u03b41, \u03b42 are the angles of lateral displacement of the tricycle\u2019s axes; \u03c1 is the radius of rotation of the tricycle\u2019s center of mass C; \u03b3 is the inclination of \u03c1 to the normal to the tricycle\u2019s longitudinal axis; Fyt is the transverse com ponent of the centrifugal inertial force relative to the tricycle\u2019s tipping axis (Fig. 1b). Consider the motion of a tricycle on a horizontal supporting surface in rotation, when a component of the centrifugal inertial force that is transverse to the tipping axis acts on the tricycle\u2019s body (Fig. 2). In Fig. 2, we show the normal reactions (Z1, Zex, Zin) and transverse reactions (Y1, Yex, Yin) of the road acting on the tricycle\u2019s single wheel and the wheels external and internal to the center of rotation. Superscripts s and n denote masses of the tricycle that are and are not subject to the springs",
            " Analogously, we may determine the distance (rela tive to the tipping axis) at which the tractional force of the mass that is not subject to the springs is applied (2) ns B/2 \u0394\u2013 h\u03bb \u03bbsin\u2013( )acos\u03b1 L ,= nn B/2 \u0394\u2013( )a \u03b1cos L .= DOI: 10.3103/S1068798X11060153 RUSSIAN ENGINEERING RESEARCH Vol. 31 No. 6 2011 BODY ROLL AND THE STABILITY OF A TRICYCLE WHEN TURNING 583 Comparing Eqs. (1) and (2), we find that (3) The transverse component of the tricycle\u2019s centrif ugal force relative to the tipping axis is (Fig. 1b) (4) where V is the linear velocity of the tricycle\u2019s longitu dinal axis. At the beginning of tipping through the front wheel and the external rear wheel, the normal reaction at the internal rear wheel is zero, and the tricycle will be stable with respect to tipping if hs + hn \u2264 G sns + Gnnn. Since hs + hn = Fyth we may now write Hence, the critical speed beyond which tipping begins is (5) where h is the height of the center of mass M; \u03b3 is found from the equation tan\u03b3 = (b \u2013 Rtan\u03b42)/R. We find that \u03b3 and (\u03b1 \u2013 \u03b3) are negligibly small"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001093_012079-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001093_012079-Figure2-1.png",
        "caption": "Figure 2. Diagram of forces acting on the rod.",
        "texts": [
            " The increased impact of the bar on the treated surface during its sliding is determined by two factors: the amount of rigidity of the bar in a static state and the speed of its rotation, which causes the appearance of a dynamic component of the force: i.e.:P = Pc + P\u0434 where P is the total force with which the rod impacts the treated surface, kg(f); Pc is the value of the component due to the statistical stiffness of the bar, kg(f); Rd is the dynamic (high-speed) component of the kg(f) force. To identify the role and approximate the value of each of these components, consider the forces impacting the rod during operation. Figure 2 shows that the following forces impact the rod: 1. The weight of the bar G = m \u00b7 g, directed vertically down. 2. The reaction of the surface Rs, which holds the outer end of the bar and deflects it from the radial direction by an angle \u03b8. 3. The force of inertia Pi, conditionally applied at the center of gravity of the bar \"C\" and tending to keep the speed of this point unchanged: Pi = m \u00b7 a. 4. The centrifugal force Pi, conditionally applied at the same point and directed along the radius of its rotation R\u05f3: RC= m \u00b7 (\u03c9\u00b4)2 \u00b7 R\u00b4, where \u03c9\u00b4 and R\u00b4, respectively, are the angular velocity and radius of rotation of point C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003338_20110828-6-it-1002.03655-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003338_20110828-6-it-1002.03655-Figure2-1.png",
        "caption": "Fig. 2. Wheeled Mobile Robot.",
        "texts": [
            " Hence, when one of the WMRs is closer than a distance threshold from an obstacle, this WMR assumes the leadership of the formation. Then it adjusts its position to track the planned trajectory. As a result of the formation control, the whole group adjusts its position as well. Thus the leadership reactively alternates so that the group may avoid collisions with the obstacles. In this section the kinematic model of the WMRs of the formation is presented. We consider that the each WMR is a differential wheeled mobile robot. The geometry of the robot is shown in Fig. 2, where (X,Y ) is the inertial coordinate system; (X \u2032 , Y \u2032 ) is the local coordinate system; Pc is the center of mass; a is the robot length; b is the distance between the actuated wheel and the axis of symmetry; r is the radius of the actuated wheel; \u03b8R and \u03b8L are the angular displacement of the right and left wheels, respectively. The WMRs of the formation present three kinematic constraints, the first imposes that the robot does not have lateral movement and the other two impose that the actuated wheels do not slip, see Coelho and Nunes (2003)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002933_s1052618810061032-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002933_s1052618810061032-Figure1-1.png",
        "caption": "Fig. 1.",
        "texts": [
            " But the Runge\u2013Kutta method of the fifth order needs significantly more time for its implementation and that is why we use the Runge\u2013Kutta method of the fourth order in what fol lows. For this purpose we present the set of equations (4), (5), in the form of the set of differential equa tions of the first order: (6) where yi \u2261 \u03bei, yi + n = , = . Further, we examine certain results of solving Eqs. (4), (5) with the following initial data: \u03b5 = 0.1, c0 = 0, \u03b10 = 0.5, \u03b11 = 2.0, \u0398 = \u03b8 = 1.4, \u03c32 = 0.01, \u03b4 = 0.5, \u0394\u03c4 = 0.1. Figure 1 depicts the plots of the dimensionless displacement of rod mean cross section (x = l/2) at the phase plane at n = 1 (Fig. 1a) and at n = 5 (Fig. 1b) and in time at n = 1 (Fig. 1c) and at n = 5 (Fig. 1d) for the case when the function \u03b1(\u03c4) is the deterministic \u03b10 + \u03b11cos\u0398\u03c4. The initial conditions are as fol lows: \u03be1(0) \u2261 \u03be01 = 1.0, (0) \u2261 = 0 at n = 1 and \u03be01 = 1.0, \u03be02 = \u2026 = \u03be05 = 0, = 0, \u2026, (0) = 0 at n = 5. For comparison to Fig. 2 the displacement variation for the same rod cross section or the imple mentation obtained when n = 1 (Figs. 2a, 2c) is shown, and for the implementation obtained at n = 5 (Figs. 2b, 2d) for the case, when the function \u03b1(\u03c4) is the stationary process \u03b1(\u03c4) = \u03b10 + \u03b11cos\u0398\u03c4 + \u03b10(\u03c4) with parameters of correlation function \u03b8 = 1",
            " Table 1 (I) contains the values of \u03bb determined for a rod subjected to periodical longitudinal force (under the deterministic definition of the problem) with different initial data and when \u03c4 = 105 for the set of equations (6) \u03be0j, ( j = 1, \u2026, n), (9) \u03b4\u03be0j, ( j = 1, \u2026, n) and under different component numbers in rod deflection expansion n. It is then seen, depending on the initial data for the undisturbed motion, that the rod can be stable or unstable. For the last case, its motion is chaotic as is seen from Fig. 1. For comparison, Table 1 (II) depicts the values of \u03bb determined for a rod subjected to the same period ical longitudinal force with applied stochastic longitudinal force under different component numbers in rod deflection expansion n for the case when by stability we mean almost certain stability. By comparison \u03b4yi' \u03b4yi n+ , \u03b4yi n+' \u20132\u03b5\u03b4yi n+ \u03c9i \u03c91 \u239d \u23a0 \u239b \u239e 2 1 \u03b1i\u2013( )\u03b4yi\u2013 \u03b2\u03b4Ii*, i\u2013 1 \u2026 n,, ,= == \u03b4\u03bei' \u03b4yj p 0( )\u2329 \u232a \u0394< \u03b4yj p \u03c4( )\u2329 \u232a < \u03b4yj 0( ) \u03b4yj p \u03c4( )\u2329 \u232a \u03c4 \u221e\u2192 lim 0= P Sup\u03c4 0> \u03b4yj \u03c4 \u03b4yj \u03c4( )\u03c4 0=,( ) \u03b4yj \u03c4( ) \u03c4 0= 0\u2192 lim 0={ } 1,= P Sup\u03c4 T\u2265 \u03b4yj \u03c4 \u03b4yj \u03c4( )\u03c4 0=,( ) 0={ } T \u221e\u2192 lim 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure11-1.png",
        "caption": "Fig. 11.Hub/Axle Displacement Analysis",
        "texts": [],
        "surrounding_texts": [
            "The authors like to acknowledge the PACE program and GM for sponsoring this project. The authors also like to thank the faculty member and students of Hongik University (South Korea) for helpful discussions in this project."
        ]
    },
    {
        "image_filename": "designv11_100_0002393_roman.2013.6628492-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002393_roman.2013.6628492-Figure2-1.png",
        "caption": "Figure 2. Smart device interface for pipeline type UV sterilizer.",
        "texts": [
            " An AVR processor based on Arduino is used for the controller. The controller processes data from UV power sensors and controls wiper when UV lamp guards get dirty. The controller provides the Bluetooth communications. 978-1-4799-0509-6/13/$31.00 \u00a92013 IEEE 352 Smart device .interface communicates with the controller through the Bluetooth. There is a \u201cSearch device\u201d button for finding pipeline type UV sterilizer. If devices are communicated each other successfully, states of UV sterilizer is shown in the monitor as Figure 2. The controller from UV sterilizer sends data as Table I. The data length is 16 bytes. It contains sterilizer ID (000~255), lamp states (On \u2013 0x01, OFF - x00), sensor states (100uW/ cm 2 ~ 400uW/ cm 2 ) and cleaning command of present cleaning state (Automatic \u2013 0x01 / Manual \u2013 0x00). Also there are some extra space for extended sensors and data. User can easily control cleaning unit with touch interface. Smart device interface provides 2 ways to control of cleaning UV lamp guard. One is automatic cleaning with timer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure23-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure23-1.png",
        "caption": "Fig. 23 Tip chamfering with EM cutting edge\u2014spiral bevel pinion",
        "texts": [
            " Spiral bevel gears have the most complex shapes in terms of tooth flank topography and blank. These gears therefore call for a general solution that is applicablewhatever the local spiral angle, face width, generating process, module or pressure angle. And again, End Mills can be used to chamfer some tooth edges, but now the options are more limited because the spiral angle can become quite significant at the Heel tooth edge for large offset hyoid pinions. For example, consider tip chamfering with the cutting edge of the End Mill as shown in Fig. 23. While chamfering the convex flank causes no issue (left, Fig. 23), chamfering the concave flank (right, Fig. 23) causes a collision risk between the tool spindle and the turn table. This approach is therefore not acceptable and the EM tool tip is to be used which, as shown in Fig. 24, causes no issue on either flank. Figure 25 shows the Tip edges of a small spiral bevel pinion chamfered using an EM. The chamfers are almost small enough to confuse with deburring. When chamfering tooth Toe and Heel edges, other considerations arise. For one, when chamfering the Toe of a pinion with a small pitch cone angle, again the turn table angle is likely to exceed the machine\u2019s limit and collide with the tool spindle (left, Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001230_0470036427.ch3-Figure3.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001230_0470036427.ch3-Figure3.3-1.png",
        "caption": "Figure 3.3 Derivation of the rms value.",
        "texts": [
            "c. value. Specifically, we would like average values of current and voltage that yield the correct amount of power when multiplied (see Section 3.4). Fortunately, such an average is readily computed: it is called the root mean square (rms) value. The rms value is derived by first squaring the entire function, then taking the average (mean), and finally taking the square root of this mean. Squaring the curve eliminates the negative values, since the square of a negative number becomes positive. Figure 3.3 illustrates this process with the curves labeled V(t) and V2(t)\u2014though we could just as well have chosen the current I(t)\u2014where the squared sinusoid, while retaining the same basic shape, is now compressed in half. If we arbitrarily label the vertical axis in units such that the amplitude Vmax \u00bc 1, it is obvious that the squared wave has the same amplitude (12 \u00bc 1). Because the squared curve resides entirely in the positive region, it is now possible to take a meaningful average. Indeed, because the curve is still perfectly symmetric, its average is simply one half the amplitude"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003705_robio.2011.6181353-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003705_robio.2011.6181353-Figure1-1.png",
        "caption": "Fig. 1. Planar model of a ball with forces at the impact points.",
        "texts": [
            " The analytic approach and formulation developed in previous work for planar impact of a single point, [1], is modified here to account for multiple point, simultaneous impact of a single body. The principles of the work-energy theorem are applied and an energetic coefficient of restitution (COR) is used for the collision to ensure energy consistency. For the purposes of this analysis, a simple test case is considered to study the ground-contact interaction of a leg\u2019s foot with its environment and to evaluate the effectiveness of the proposed approach. Thus, a planar model of a ball which undergoes multiple impact points is examined. The ball, shown in Fig. 1 with radius R, has three degreesof-freedom (DOFs) denoted by generalized coordinates q1 and q2 which are translational and q3 which is rotational. The ball\u2019s position is indicated by the position vector PNO. At the point of impact, the ball is in simultaneous contact with the ground, point 1, and the wall, point 2. Two force components are present at each impact point, as shown in Fig. 1. One force is normal to both surfaces and the other is tangential to both surfaces, caused by friction. A. Rodriguez is with the Department of Mechanical and Aerospace Engineering, University of Texas at Arlington, Arlington, TX 76017 USA adrianrodriguez2009@mavs.uta.edu A. Bowling is with the Department of Mechanical and Aerospace Engineering, University of Texas at Arlington, Arlington, TX 76017, USA bowling@uta.edu The remainder of the paper is organized as follows. Background will be presented on the methods used to treat the rigid body collision considered in this work",
            " The equations of motion are revisited and a definite integration of (2) over a short time \u03b5 for the impact event,\u222b t+\u03b5 t (A q\u0308+ g (q)) dt = \u222b t+\u03b5 t JT (q) F dt (8) gives, A ( q\u0307(t+ \u03b5)\u2212 q\u0307(t) ) = JTp (9) It should be noted that for the modeling of a legged robot, which consists of a multi-body system, internal forces and torques would be present in the equations of motion. Although, these forces would be considered very small and negligible compared to the magnitude of the impulse forces after integration of the equations over the short impact event. Hence, it is appropriate to continue this analysis for the test case considered in Fig. 1. Multiplying the inverse of the mass matrix on each side of (9) and using q\u0307 to represent the change between preand post-impact generalized speeds, q\u0307 = A\u22121JTp = L p = [Lt1 | Ln1 | Lt2 | Ln2] p (10) where L \u2208 3X4 can be termed as a collision matrix. The columns of L are shown to uncover their respective normal and tangential contributions as column vectors. Finally, the product can be carried out such that the equations of motion take the form of, q\u0307 = Lt1 pt1 + Ln1 pn1 + Lt2 pt2 + Ln2 pn2 (11) In this work, Coulomb friction is used to characterize the relationship between normal and tangential forces using a coefficient of friction \u03bc, ft = \u00b1\u03bc |fn| (12) where the sign in (12) will depend on the direction of friction at the impact point. Given that the method used here considers the impulses at impact, (12) can be expressed in terms of impulses as in [16], [17] after a definite integration, pt = \u00b1\u03bc |pn| (13) An expression for the tangential impulses at each impact point using Coulomb\u2019s friction can be written for the model illustrated in Fig. 1 as, pt1 = \u00b1\u03bc1 |pn1| pt2 = \u00b1\u03bc2 |pn2| (14) The point at which no-slip occurs at the ground and wall impact points can be indicated by vt1 = 0 and vt2 = 0, respectively. Using Coulomb friction, the tangential impulses for no-slip can be expressed as, pt1 = m1 |pn1| pt2 = m2 |pn2| (15) where m1,m2 \u2264 \u03bcs. The terms m1, m2 are coefficients of friction for no-slip and \u03bcs is the static coefficient of friction. The specific model considered in Fig. 1 will be treated with the theory developed here, which uses (11) and (14) without any loss of generality. The absolute value of the normal impulse terms in (11) can be taken while preserving their signs and the expressions in (14) can be substituted into (11) to give, q\u0307 = Lt1(\u00b1\u03bc1|pn1|) + Ln1(|pn1|) +Lt2(\u00b1\u03bc2|pn2|) + Ln2(\u2212|pn2|) (16) where the sign of the tangential terms are kept ambiguous and will depend on the direction of friction acting at each impact point. The two normal impulse terms in (16) can be collected to obtain, q\u0307 = (Ln1 \u00b1 \u03bc1Lt1) |pn1|+ (\u2212Ln2 \u00b1 \u03bc2Lt2) |pn2| (17) Furthermore, the constraint equation obtained in (7) can be rewritten to include the Coulomb friction expressions in (14) and the magnitudes of the normal impulse terms",
            " In the case of noslip, then the coefficient of friction for a tangential velocity of zero, mi \u2208 i = 1, 2 is used in (16). The change of signs in (19) cause a shift in the Wn vs. |pn1| plot and updated values for the normal impulse contributions must be calculated for the remaining velocity components. After |pnc| is reached in the colllision, then the net normal work, Wnf , can be calculated using the energetic COR. Lastly, the end condition for the normal impulse, |pnf |, is evaluated to allow for the calculation of the post impact velocities using (31). Here, a simulation of a ball impacting a corner, as in Fig. 1, is performed for a planar case to show that the proposed method yields an analytic and energetically consistent solution for the post impact velocities. A static and dynamic coefficient of friction of 0.5 and 0.25 is used for the impacting surfaces. The initial, pre-impact, and post impact states of the ball are presented in Table I. The simulation was started with some initial conditions and the remaining states were determined from the application of the algorithm discussed in Sec. V for the rigid body collision"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003093_amm.204-208.4518-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003093_amm.204-208.4518-Figure2-1.png",
        "caption": "Fig 2 Configuration of thruster system",
        "texts": [
            " The target of allocation was the optimization of the power consumption, for improving the economical efficiency of the system; the target function being: 2 2 2 1 2 1 ( ) ( ) N i i i f x x x \u2212 = = +\u2211 (5) The confined conditions make sure that the thrust outputs of the thrusters equal to the thrust calculated by the controller, namely, 1 2 1 1 2 2 1 3 2 1 , 2 , 1 1 ( ) 0 ( ) 0 ( ) 0 N treq i i N treq i i N N treq i y i i x i i i g x X x g x Y x g x N x l x l \u2212 = = \u2212 = = = \u2212 = = \u2212 = = \u2212 + =          \u2211 \u2211 \u2211 \u2211 (6) And make sure that the thrust allocated would not exceed the maximum thrust output provide by the thrusters, namely, 2 2 4 max,1 1 2 2 2 3 max, 2 1 2 ( ) 0 ( ) 0 N N N N g x T x x g x T x x+ \u2212 = \u2212 + \u2265 = \u2212 + \u2265      (7) As azimuth thrusters were used, the dead zone must be taken into consideration. Consequently, the limitation of the angle of thrust are introduced in the confined conditions. The configuration of thruster system of the platform, see Fig 2. When simulation was carried out, the following steps were executed: first, platform state was measured, and the state signal was filtered by Kalman filter. Then, the state filtered would be contrast with the setting state, and the error would be sent to the PID controller. After that, the force should be applied to the platform to revert to the setting point was obtained by the controller, and wind load was added to the force. Afterward, the complex force was allocated by means of distribution logics, getting the thrust of each thruster"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure15-1.png",
        "caption": "Figure 15. Force of Shoulder Muscle",
        "texts": [
            " Input to Articulatio Acromioclavicularis (Analysis 1) External input is added at the articulation sternoclavicularis (Fig. 12 and Table V). a F represents arm\u2019s weight. b F represents the external input. The force of shoulder muscle, and the force of clavicle and the rib are shown in Fig. 13. B. Input to Scapulothoracic Joint (Analysis 2) External input c F is added at the scapulothoracic joint (Fig. 14 and Table VI). The force of shoulder muscle, the force of clavicle and the rib are shown in Fig. 15. C. Input to Articulatio Acromioclavicularis and Scapulothoracic Joint (Analysis 3) External input is added evenly at the articulation sternoclavicularis ( b F ) and the scapulothoracic joint ( c F ) (Fig. 16 and Table VII). The force of clavicle and the rib are shown in Fig. 17. D. Discussion In order to minimize the burden on the shoulder muscles, we compared the amount of force exercised by the muscles in the analysis 1 to 3 by integrating the absolute values of muscles forces 1 F , 2 F , 3 F , 4 F , when 2 \u03b8 is varied from 80 degrees to 110 degrees, as this is supposed to be the range most used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002459_urai.2012.6463047-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002459_urai.2012.6463047-Figure2-1.png",
        "caption": "Fig. 2. Reader antenna fixed type RFID tag reading system",
        "texts": [
            " Attached position data of the RFID tags to the workpiece is stored in the RFID tag as the workpiece information, and used for the measurement. We propose two realization methods for the measurement of workpiece position and posture by using RFID tags: reader antenna fixed type and moved type RFID tag reading system. 2.1 Reader Antenna Fixed type Tag Reading System In the reader antenna fixed type RFID tag reading system, reader antennas are arranged in a matrix pattern under the workpiece tray (Fig. 2.). These fixed antennas communicate the tags attached to the workpiece, and the workpiece position and posture are measured. This system can read the data in the tags instantly after the workpieces are put on the tray. 2.2 Reader Antenna Moved type Tag Reading System In the reader antenna moved type RFID tag reading system, the data in tags are read by relative motion of the RFID tag and the tag reader (Fig. 3.). Workpiece position and posture are measured by the start and stop time of tag reading and the speed of the reader movement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001736_978-3-319-01851-5_12-Figure12.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001736_978-3-319-01851-5_12-Figure12.8-1.png",
        "caption": "Fig. 12.8 A view of the three-wheeled robot with Mekanum wheels in a",
        "texts": [
            " On the other hand, the determination of the motion produced by a given history of joint torques requires (a) the calculation of I, which can be achieved symbolically; (b) the inversion of I, which can be done symbolically because this is a 2 2 matrix; (c) the calculation of the Coriolis and centrifugal terms, as well as the dissipative forces; and (d) the integration of the initial-value problem resulting once initial values to and P a have been assigned. We now consider a three-dof robot with three actuated wheels of the Mekanum type, as shown in Fig. 10.19, with the configuration of Fig. 12.7, which will be termed, henceforth, the -array. This system is illustrated in Fig. 12.8. Below we will adopt the notation of Sect. 10.5.2, with \u02db D =2 and n D 3. We now recall that the twist of the platform was represented in planar form as t0 ! Pc (12.55) where ! is the scalar angular velocity of the platform and Pc is the two-dimensional position vector of its mass center, which will be assumed to coincide with the -array centroid of the set of points fCi g31. Moreover, the three wheels are actuated, and hence, the three-dimensional vector of actuated joint rates is defined as P a 2 4 P 1P 2P 3 3 5 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.111-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.111-1.png",
        "caption": "Fig. 2.111 4PRRRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology P\\R||R\\R||Pa",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002718_amm.197.46-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002718_amm.197.46-Figure3-1.png",
        "caption": "Figure 3. The modal shapes of the positioning platform",
        "texts": [
            " Release Y axial DOF and couple the DOF in other directions on the surface which and couple the DOF in other directions on the surface which is between front shafting bearing and front lapping plate, back shafting bearing and back lapping plate. As a result, the finite element model of the positioning platform with connecting plate is established, and it is shown in Fig. 2. The natural frequencies of the XY positioning platform are gained through modal analysis, and they are summarized in Table 1. The first 6 th modal shapes of the positioning platform are shown in Figure 3. As shown in Fig.3, at the frequency of 165Hz, the X axial and Y axial tables driven by X axial motor and Y axial connecting plate driven by Y axial motor generate torsion vibration around the X axis in the flexure hinges of the elastic decoupling mechanism.. At the frequency of 184Hz, X axial and Y axial positioning platform and the Y axial connecting plate generate torsion vibration around the X axis in the flexure hinges of the elastic decoupling mechanism, while a torsion vibration around the Z axis at the same time"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002837_2011-01-0197-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002837_2011-01-0197-Figure1-1.png",
        "caption": "Figure 1. Quarter of conrod",
        "texts": [
            " For example, the bearing and the conrod form a single body. The finite element size in the bore of the conrod is around 5 mm. Generally, just one distorted element (1.5 mm thick 5 mm wide) is used to represent the bearing thickness. With linear elastic behavior of the bearing it is impossible to predict realistic stresses in the substrate and in the overlay of the bearing. The major and realistic result obtained from the Elasto-Hydrodynamic simulation is the fluid pressure acting in the crown area of the conrod as illustrated on figure 1. Experience from the field has pushed bearing suppliers to propose different materials and structures. The most common type of bearing are given on figure 2 & 3. The most used bearing (figure 2), called bimetallic bearing, is mainly made of a steel back and an antifriction substrate. An intermediate layer (40 \u00b5m thick) creates the bonding between these two parts. The second type of bearing (figure 3), or trimetallic bearing, is composed of three parts: the bearing back in steel (sometime in copper alloy), the substrate, and the overlay"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002894_amr.837.88-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002894_amr.837.88-Figure1-1.png",
        "caption": "Fig. 1",
        "texts": [
            " For the case of general method - the method multibody - the exact method are established liaison relationships between the parameters iii yx \u03b8,, , write matrices ][],[ CB , ][M inertia matrix. Use Lagrange equations, if non-holonomic constraints. Matrix differential equation of motion is written and it can be solved numerically using RungeKutta method of order four. Of the iterative method, we obtain the parameters used in calculating the reaction force N expression that can be evaluated accurately in journal bearings behaviour. Any would be their source of appearance, they usually produce unwished effects during the mechanisms\u2019 functioning. The model presented in Figure. 1 can be utilized to study the influence of the clearance in the case of great speeds, too. 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: 128.119.168.112, Univ of Massachusetts Library, Amherst, USA-10/07/15,09:40:47) To this goal, we consider the motor element has a constant angular speed \u03c9 , realize the cinematic analysis in the hypothesis 0=j with the aid of the relations: 112 11 sin 2 ;cos 2 coscos;sinarcsin; \u03b8\u03b8 \u03b8\u03b8\u03b8\u03b8\u03c9\u03b8 l y l xx lbOBx l b t cc B =\u2212= +==      == "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003687_detc2011-47626-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003687_detc2011-47626-Figure6-1.png",
        "caption": "Fig. 6 Thermal expansion of bearing components",
        "texts": [
            "0468 / PruN          , where lengthC is the length of the convective fluid, and Re is the Reynolds number of the cooling air flow, Re /L U     . Pr is the Prandtl number, Pr /p convC k  . At the housing-ambient air interface, the heat convection is solved with the following equation.   _ _ 0, , ho h conv h conv h h x y ho r R T k h T R T r         , (10) where _conv hh is the heat transfer coefficient between the housing and the ambient air. The bearing components have different thermal expansions due to the temperature rise of the foil bearing, which will lead to a change in the radial clearance, as shown in Fig. 6. Terms s , b , and h represent the thermal expansion of the shaft, the bump foil, and the housing. The variation of the radial clearance due to the thermal growth can be given as s b hC       (11) The shaft and the housing are assumed to have free expansion, which is linear with the temperature rise, hi s h hi h r R h s s r R s R T R T             , (12) where s denotes the thermal expansion coefficients of the shaft, and h is that of the housing. The thermal expansion of the bump foil is shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003110_s1068798x11060153-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003110_s1068798x11060153-Figure3-1.png",
        "caption": "Fig. 3. Forces on a tricycle as it turns on a banked road.",
        "texts": [
            "= Assuming a linear relation between the lateral force on the wheel and the tire\u2019s angle of lateral displace ment, we find that [2] where k1, k2 characterize the tires\u2019 resistance to lateral displacement at the front and rear wheels; \u03c8 is the road\u2019s drag coefficient. Since the tricycle\u2019s inner wheel breaks away from the road surface at the beginning of tipping, the angle \u03b42 of lateral displacement increases because all the transverse load applied to the rear wheels acts on the outer wheel in that case. This must be taken into account when using Eq. (5) or Eq. (6). Equation (6) assumes motion over a flat horizontal supporting sur face. In the general case of turning on a banked road (Fig. 3), on the assumption that the wheel is not inclined with respect to the frame, tipping will not occur if (7) Substituting Eq. (3) into Eq. (7), we obtain We know that hs + hn = Fyth, + = Fyt, G shs + Gnhn = Gh, G s + Gn = G. \u03b41 mb k1L V 2 R \u03c8g \u03b8sin+\u239d \u23a0 \u239b \u239e ; \u03b42 mV 2a 2k2LR ,= = Fyt s hs \u03b2cos Fyt n hn \u03b2cos+ Fyt s ns \u03b2sin Fyt n nn \u03b2sin+\u2264 + G sns \u03b2cos G nnn \u03b2cos G shs \u03b2sin G nhn \u03b2.sin+ + + Fyt s hs Fyt n hn+( ) \u03b2cos Fyt s Fyt n+( )nn \u03b2sin Fyt s ah\u03bb \u03bbsin\u2013\u2264 \u00d7 \u03b1cos \u03b2/Lsin Gs Gn+( )nn \u03b2cos G sah\u03bb \u03bbsin \u03b1cos\u2013+ \u00d7 \u03b2/Lcos Gshs Gnhn+( ) \u03b2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003610_imece2013-63166-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003610_imece2013-63166-Figure4-1.png",
        "caption": "FIGURE 4: SKETCH OF (a) MOBILE PLATFORM; (b) A LEG",
        "texts": [
            " Authors like [9] studied the inverse dynamics of a 3-DOF spatial PKM using the Newton-Euler approach. The expression used to obtain the reactions in the links is, Eqn. (13): Mstatic \u00b7qdin = fdin (13) where Mstatic is the matrix of the statics, qdin is the vector with the reactions in the rotational and prismatic joints and fdin is the vector with all the forces of the manipulator, i.e., the applied forces (if any exist), the self-weight and the inertial forces. To obtain the Mstatic, both the static analysis of the mobile platform and the static analysis of the links are developed. In Fig. 4a a skecht of the mobile platform and a leg of the manipulator with the current notation are shown. The sum of applied forces, reactions and moments of the mobile platform referred to the fixed frame is given by Eqn. (14), and for each leg by Eqn. (15). [ RA1 rA1 \u00d7RA1 ] + [ RA2 rA2 \u00d7RA2 ] + [ RA3 rA3 \u00d7RA3 ] + [ Fe re \u00d7Fe ] = [ 0 0 ] (14) [ \u2212RAi \u2212rAi \u00d7RAi ] + [ RBi rBi \u00d7RBi ] + [ 0 MBi ] + [ FEi rei \u00d7FEi ] = [ 0 0 ] (15) A system of 24 unknowns and 24 equations is obtained and the Mstatic can be defined as the matrix with the coefficients that multiplies the reactions in the static problem"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003395_1.4816689-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003395_1.4816689-Figure8-1.png",
        "caption": "Fig. 8. The safety envelope for a spherical shot translating and rising at burst speed, with detail (inset): xs\u00bc 0, ys\u00bc 10 m, R\u00bc 1 m, vs,x\u00bc 5 m/s, vs,y\u00bc 5 m/s, u\u00bc 5 m/s.",
        "texts": [
            " (10) and(11) become x1 \u00bc xs R if 0 s s1; xs \u00fe \u00f0wu u\u00de w2 u \u00fe u2 R g s\u00fe 2 w2 u w2 u \u00fe u2 us if s > s1; 8>< >: (56) y1 \u00bc ys \u00fe us 1 2 gs2 if 0 s s1; ys \u00fe 2 wu w2 u \u00fe u2 uR 1 2 \u00f0w2 u u\u00f02gs u\u00de\u00de w2 u \u00fe u2 gs2 if s > s1; 8>>< >>: (57) x2 \u00bc xs \u00fe 2 \u00f0wu u\u00de w2 u \u00fe u2 R g s 2 w2 u w2 u \u00fe u2 us if 0 s s1; xs R if s > s1; 8>< >: (58) y2 \u00bc ys \u00fe 2 wu w2 u \u00fe u2 u R 1 2 \u00f0w2 u u\u00f02gs u\u00de\u00de w2 u \u00fe u2 gs2 if 0 s s1; ys \u00fe us 1 2 gs2 if s > s1: 8>>< >>: (59) 759 Am. J. Phys., Vol. 81, No. 10, October 2013 G. R. Heppler and G. M. T. D\u2019Eleuterio 759 Downloaded 21 Sep 2013 to 152.11.242.100. Redistribution subject to AAPT license or copyright; see http://ajp.aapt.org/authors/copyright_permission Not unexpectedly, this case has a similar piecewise form to the previous one. The envelope and trajectories are shown in Fig. 8, where there is a vertical boundary line at x\u00bc x2\u00bc xs R (shown as a heavy solid line and corresponding to s s1) as well as another bounding curve above and to the right (shown as a heavy dashed line) that corresponds to (x1, y1) for s s1. There is a tangency envelope interior to the safety envelope. This part of the envelope arises when Eqs. (56) through (59) are evaluated on 0 s s1. It is shown in more detail in the inset of Fig. 8 where we observe that the interior portion of the envelope curve is present to satisfy the tangency condition. When s\u00bc 0, x1\u00f00\u00de \u00bc xs R; x2\u00f00\u00de \u00bc xs; (60) y1\u00f00\u00de \u00bc ys; y2\u00f00\u00de \u00bc ys R; (61) and, at s1\u00bc u/g, x1;2\u00f0s1\u00de \u00bc xs R; (62) y1;2\u00f0s1\u00de \u00bc xs \u00fe 1 2 u2 g : (63) The parameter cases considered in Secs. III A through III G illustrate the range of envelopes that can occur for different choices of the system parameters vs,x, vs,y, u, and R. All of the results arise from the general envelope expressions in Eqs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.29-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.29-1.png",
        "caption": "Fig. 2.29 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PRPPa (a) and 4PRPaP (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology P\\R||P\\Pa (a) and P\\R\\Pa \\kP (b)",
        "texts": [
            "21b) The three first revolute joints of the four limbs have parallel axes 3. 4PRPaR (Fig. 2.27a) P\\R\\Pa\\kR (Fig. 2.21c) The second and the last joints of the four limbs have parallel axes 4. 4PPPaR (Fig. 2.27b) P\\P\\kPa\\R (Fig. 2.21d) The last revolute joints of the four limbs have parallel axes 5. 4PRPPa (Fig. 2.28a) P||R\\P||Pa (Fig. 2.21e) The second joints of the four limbs have parallel axes 6. 4PPRPa (Fig. 2.28b) P\\P||R\\Pa (Fig. 2.21f) The third joints of the four limbs have parallel axes 7. 4PRPPa (Fig. 2.29a) P\\R||P\\Pa (Fig. 2.21g) Idem No. 5 8. 4PRPaP (Fig. 2.29b) P\\R\\Pa \\kP (Fig. 2.21h) Idem No. 5 9. 4PPPaR (Fig. 2.30a) P\\P\\kPa\\kR (Fig. 2.21i) Idem No. 4 10. 4PPPaR (Fig. 2.30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003623_pat.1870-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003623_pat.1870-Figure3-1.png",
        "caption": "Figure 3. Principles of the electro-optical effects after non-contact photo-alignment process. (a) Dark state before applied voltage. (b) Bright state after applied voltage.",
        "texts": [],
        "surrounding_texts": [
            "Conversion percentage of the NLC-B1 and NLC-B2 mixture systems We placed NLC-B1 and NLC-B2 mixture solutions with specific concentrations under UV light exposure. During the UV light Figure 2. Chemical structures of the components used for the formation of the vertical alignment composite film. wileyonlinelibrary.com/journal/pat Copyright 2011 John Wiley & Sons, Ltd. Polym. Adv. Technol. 2012, 23 299\u2013310 3 0 2 exposure, free radicals would be produced by the photoinitiator, and subsequently attack the unsaturated acetyl double bonds, leading to copolymerization between the acrylic monomers A and B and thus lowering the LC solubility. Phase separation would then occur, resulting in the formation of layered structures. The simultaneous destruction of the acetyl double bonds could also lead to a reduction of the UV absorption with increase in the UV light exposure time until the reactions terminated and the saturation value was attained.[70] Therefore, we measured the UV absorption wavelength range and maximum absorption wavelength (Abs, lmax) with respect to the exposure time for the two LC mixture systems, after UV light exposure for 0\u201330min with a UV\u2013Vis absorption spectrometer. From the CP versus UV light exposure time plots (Fig. 4), one can clearly observe that the CP value had already reached 89% when the NLC-B2 mixture system was placed under UV light exposure for about 10min. Its polymerization reaction rate far exceeded that of NLC-B1, and the CP value after 30min was already close to 99%. From the foregoing result, one can infer that under the same UV light exposure condition, the degree of copolymerization reaction was more complete for NLC-B2 than for NLC-B1. The reason we selected two LC mixture systems whose CP values were 89% was that with insufficient UV light exposure time, the phase separation process between the LC molecules and photo-curable acrylic monomer would be incomplete, leaving a portion of the LC molecules \u2018\u2018wrapped\u2019\u2019 within the polymer to form a PDLC film structure. On the other hand, when the UV light exposure time was exceedingly long, although the phase separation process would be nearly complete, the crosslinking density would be larger and the degree of crystallization would increase rapidly, which would likely lead to drawbacks such as brittleness of the VACOF, etc.[69,70] In order to achieve uniformity in the experiment conditions and avoid occurrence of the foregoing conditions, we therefore selected NLC-B1 and NLC-B2 mixture systems with approximately the same CP values, to proceed with the related experiment processes."
        ]
    },
    {
        "image_filename": "designv11_100_0002407_amr.189-193.1409-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002407_amr.189-193.1409-Figure2-1.png",
        "caption": "Fig. 2. The UPU limb which afford cconstraint 1f",
        "texts": [
            " Since the limb constraint system consist one constraint force 1f , the twists of limb can be obtained: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= , 1 4 4 4[0,0,0; , , ]P d e f= , 2 5 5 5[0,0,0; , , ]P d e f= The new revolute joints 4 5,R R can be obtained by the linear combination of revolute joint and prismatic joint, so the twist system of limb could become: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= 4 1 1 1 4 4 4[ , , ; , , ]R a b c d e f= , 5 2 2 2 5 5 5[ , , ; , , ]R a b c d e f= Therefore, the UPU limb can be obtained to afford the constraint force 1f , as shown in Fig. 2. In the similar method, the SPS limb can be obtained to afford the constraint no force, as shown in Fig. 3. Finally, The symmetrical decoupled SPM (Fig. 4-a) and non-symmetrical decoupled SPM(Fig. 4-b) can be obtained by assembling the limbs synthesized perpendicularly. Since each UPU limb affords one constraints f1, the 3 UPU\u2212 parallel mechanism (Fig. 4-a) and 2SPS RU\u2212 parallel mechanism (Fig. 4-b) is non-over-constrained. The principle of decoupled motion of UPU RU SPS\u2212 \u2212 SPM are as following: when the active prismatic joint of UPU chain works, the moving-platform of spherical parallel mechanism rotates around axis X (actuated by the motor 1), while the actuators of two kinematic chains holds as the kinematic chain has the kinematic joint along axis X"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003676_icems.2013.6754494-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003676_icems.2013.6754494-Figure1-1.png",
        "caption": "Fig 1. Cross section of spoke type UFDMPM",
        "texts": [
            " This special type of DMPM has more advantages than the traditional DMPM [2][3], which are named separated field DMPM or un-unified field DMPM for distinction and can be applied in HEV. The most prominent advantage is the torque strengthening effect which benefited from its unique electromagnetic coupling characteristic, and this effect has been researched in reference [3]. If negative daxis current is applied to stator of UFDMPM, the torque output of inner rotor will be strengthened by the electromagnetic coupling effect, and vice versa. Fig 1 is the cross section of the spoke type Unified Field Dual Mechanical Port Electric Machine. Its voltage equation, flux equation and torque equation are respectively expressed as equations (1)-(3), which are derived in reference [5]. 0 0 00 0 0 0 0 00 0 0 0 0 00 0 0 0 0 00 0 0 do do do dos qo qo qo qos di di di diir qi qi qi qiir u iR u iR d u iR dt u iR \u03c8 \u03c8\u03c9 \u03c8 \u03c8\u03c9 \u03c8 \u03c8\u03c9 \u03c8 \u03c8\u03c9 \u2212\u23a1 \u23a4 \u23a1 \u23a4 \u23a1 \u23a4 \u23a1 \u23a4\u23a1 \u23a4\u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u23a2 \u23a5\u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u23a2 \u23a5\u23a2 \u23a5= + + \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u23a2 \u23a5\u23a2 \u23a5 \u2212 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u23a2 \u23a5\u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u23a3 \u23a6 \u23a3 \u23a6\u23a3 \u23a6 \u23a3 \u23a6 \u23a3 \u23a6 \u23a3 \u23a6 (1) 0 0 0 0 0 0 0 0 0 0 0 0 do do ddio do mo qo qo qo di ddoi di di mi qi qi qi L M i L i M L i L i \u03c8 \u03c8 \u03c8 \u03c8 \u03c8 \u03c8 \u2212\u23a1 \u23a4 \u23a1 \u23a4 \u23a1 \u23a4 \u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5= + \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u2212 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a3 \u23a6\u23a3 \u23a6 \u23a3 \u23a6 \u23a3 \u23a6 (2) ( ) ( ) ( ) ( ) 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002695_0954406213507705-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002695_0954406213507705-Figure1-1.png",
        "caption": "Figure 1. Schematic view of a Stewart platform. (a) Side view; (b) vertical view.",
        "texts": [
            " \u2018\u2018Dynamic modeling\u2019\u2019 section formulates the dynamic modeling of the uncertain Stewart platform, including the multi-body dynamics, platform friction, PMSM dynamics, and the actuator faults. The controller design, stability analysis, and tracking error analysis of the novel control and the MBC schemes are derived in \u2018\u2018Control design\u2019\u2019 section. In \u2018\u2018Simulation results\u2019\u2019 section, the comparisons are analyzed in the fault-free and faulty conditions. Finally, conclusions are drawn in \u2018\u2018Conclusions\u2019\u2019 section. Multi-body dynamics of the Stewart platform The primary structures of the Stewart platform (Figure 1) are one fixed base, one movable platform, and six extendable actuators, where Ob XbYbZb, OP XPYPZP, da, db, ra, rb, and l0 denote the reference frame, platform frame, upper joint spacing, lower joint spacing, upper joint radius, lower joint radius, and middle actuator length, respectively. Considering the influences of the motor systems, actuator inertial forces and platform friction,36 the multi-body dynamics of the Stewart platform can be formulated by17 JTy,q v\u00f0 \u00deTa \u00bcMs,m v\u00f0 \u00de \u20acq\u00fe Cs,m v, _q\u00f0 \u00de _q \u00fe Gs,m v\u00f0 \u00de \u00fe JTy,q v\u00f0 \u00deffr\u00f0xm\u00de \u00f01\u00de ffr\u00f0xm\u00de \u00bc fv\u00f0xm\u00de \u00fe fc\u00f0xm\u00de \u00fe fs \u00f02\u00de fv, i\u00f0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003793_20130708-3-cn-2036.00014-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003793_20130708-3-cn-2036.00014-Figure1-1.png",
        "caption": "Fig. 1. The trajectories of the leader and the 10 followers for 29s. 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 lines are the trajectories of 10 followers, respectively. The asterisks are the original locations of the 10 followers.",
        "texts": [
            " In the proof, we magnified V\u0307 (e(t)) and assumed that the functions \u2016ei(t)\u2212 e j(t)\u2016exp(\u2212\u2016ei(t)\u2212e j(t)\u20162 c ) are at their maximum values for all i and j. Thus, the actual value of the error is much smaller than \u03b5 . In this section, a numerical example is shown to illustrate the above results. In the simulation, the system consists of ten followers and a leader with dynamics described by (1) and (2). We describe the dynamics of the leader of followers in a twodimensional Cartesian coordinate system with g(y) =\u2212y ( 1\u221220exp(\u2212\u2016y\u2016 2 0.2 ) ) . Fig. 1 shows the trajectories of the leader and the 10 followers, according to the model (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]. Fig. 2 presents the errors 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 is as follows. \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 0 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002653_amr.189-193.2037-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002653_amr.189-193.2037-Figure1-1.png",
        "caption": "Fig. 1 The curvilinear shaped teeth of the face-gear. Fig. 2 Schematic illustration of the head-cutter and the face-gear.",
        "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. (ID: 130.113.86.233, McMaster University, Hamilton, Canada-01/04/15,18:37:20) The contents of the paper cover: (1) Concept of generation of face-gear. (2) The machining mathematical model of face-gear. (3) Simulation of cutting. Basic Idea for Generation of Face-gear The face-gear with curvilinear shaped teeth is shown in Fig.1. As illustrated in Fig.1, two coordinate frames are defined; o and ),(0 VHo where o is defined on the pitch plane, ),(0 VHo is the set at the center of the head-cutter. The tilt of the head-cutter allows to avoid interference of the hear-cutter with teeth that neighbor to the space being generated (Fig.2). The gear tooth generation process is based on the following kinematics: the head-cutter performs rotational motion (\u2170) about its inclined axis and the gear being generated is rotated about its axis (\u2171) while the imaginary rack-cutter is provided with translational motion (\u2172), in order to obtain the rolling motion required for tooth form generation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003854_20120905-3-hr-2030.00044-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003854_20120905-3-hr-2030.00044-Figure12-1.png",
        "caption": "Fig. 12. 2-link telescoping biped robot model",
        "texts": [
            " We define that L1 (m) is a length of the virtual stance leg, L2 (m) is a length of the virtual swing leg, \u03b31 (rad) is an angle of the virtual stance leg, \u03b32 (rad) is an angle of the virtual swing leg, \u03b1 = \u03b84 \u2212 \u03b83 (rad) is the relative angle between the shank and foot of the swing leg and F (\u03b1) (N) is a push-off force as shown in Fig. 11. For simplicity, we then assume l1 \u2248 0 (m), l2 \u2248 0 (m), l3 \u2248 1.0 (m), m1 \u2248 0 (kg), m2 \u2248 0 (kg), m3 \u2248 5.0 (kg), D \u2248 0 (Nm/(rad/s)) and \u03b2 \u2248 0 (Nm/(rad/s2)). When these assumptions are almost satisfied, we can notice L1 = L2 = L and \u03b31 = \u03b32 = \u03b3 as shown in Fig. 12. We can, moreover, infer that a linear relationship exists between \u03b1 and F (\u03b1) since the push-off force substantially depends on the restoring torque due to the ankle elasticity of the swing leg. We thus assume that the push-off force is F (\u03b1) = \u2212\u03b5\u03b1 and the direction of F (\u03b1) is equal to the direction from the hip of the swing leg to the tip of the swing leg. Using these assumptions, we can deal with the complicated biped robot as the 2-link biped robot with telescopic legs as shown in Fig. 12. We therefore transform this push-off force into the torque around the contact point of the stance leg as shown in Asano and Luo (2008). We first derive a Jacobian matrix from the end of the stance leg to the end of the swing leg. Using this geometric relationships as shown in Fig.12, we obtain J = [ Lcos(\u03b3) \u2212Lcos(\u03b3) Lsin(\u03b3) Lsin(\u03b3) ] . (13) Transforming the push-off force using (13), we can get[ \u03c4ankle \u03c4hip ] = JT [ sin(\u03b3) cos(\u03b3) ] F (\u03b1) = [ \u2212\u03b5\u03b1Lsin(2\u03b3) 0 ] , (14) where \u03c4ankle is transformed torque around the contact point of the stance leg and \u03c4hip is transformed torque around the hip. Since \u03b1 and \u03b31 monotonically increase during the DSP, we assume a relationship between \u03b1 and \u03b31 as \u03b1 = \u03b7\u03b31 + \u03b6, (15) where \u03b7 and \u03b6 are constants and \u03b7 is positive. Substituting (15) into (14), we get \u03c4ankle = \u2212\u03b5(\u03b7\u03b31 \u2212 \u03b6)Lsin(2\u03b3), = \u2212\u00b5(\u03b31 \u2212 \u03bd), (16) where \u00b5 = \u03b5\u03b7Lsin(2\u03b3) and \u03bd = \u03b6/\u03b7"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002567_icmtma.2011.343-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002567_icmtma.2011.343-Figure6-1.png",
        "caption": "Figure 6. The model had restricted and press",
        "texts": [
            " So we constrained all degrees of freedom on the profile of the wire race. For the ball, it has not only plastic deformation but also displacement relative to wire race. To ensure that the model is reasonable, for the two sections, by the symmetry of the model, we should constrain it\u2019s normal displacement having no any influences on tangential displacement. By the given load of platform, we applied surface load which is perpendicular to the ball section on 1/4 spherical. The value is 0.6N/ mm 2 .The model is shown in Figure 6. The equivalent stress map is shown in Figure 7 (a) and (b). Figure 7(b) is the local magnification map on contact area. Figure 8 shows the equivalent displacement map. We can see that the maximal contact stress of contact region is 100Mpa, the maximal displacement is 0.009mm and the contact stress of wire is significantly larger than that of ball. (1) In the course of bearing\u2019s practical work, wires are often damaged erlier than balls. The reason is that the yield limit and ultimate strength of 65Mn quenched spring steel are 800MPa and 1000MPa respectively, the minimum allowable safety factor is about 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001154_978-3-030-60986-3-Figure4.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001154_978-3-030-60986-3-Figure4.10-1.png",
        "caption": "Figure 4.10 The circuit of solution of problem 4.11",
        "texts": [
            " The equivalent inductance of the circuit, in the second test, is: Leq 2 \u00bc L1 \u00fe L2 2M \u00f04\u00de Herein, minus sign is applied for the mutual inductance because of the inverse position of the dots of the inductors. Solving (3) and (4): Leq 1 \u00fe Leq 2 \u00bc 2 L1 \u00fe L2\u00f0 \u00de \u00f05\u00de Solving (1), (2), and (5): 8 mH \u00bc 2 L1 \u00fe L2\u00f0 \u00de ) L1 \u00fe L2 \u00bc 4 mH \u00f06\u00de Solving (1), (3), and (6): 6 mH \u00bc 4 mH \u00fe 2M ) M \u00bc 1 mH Choice (4) is the answer. 4.11. The positions of the dots of the mutually coupled coils can be determined by using the right-hand rule. As is shown in Figure 4.10.1, the magnetic fluxes of the coils oppose each other, for an arbitrary direction of the current flowing through the coils. Therefore, the positions of the dots of the coils must be like the ones illustrated in Figure 4.10.2. The equivalent inductance of the mutually coupled coils can be calculated as follows: Leq \u00bc 1\u00fe 2 2 0:5 \u00bc 2 H \u00f01\u00de Based on the information given in the problem, \u03c9 \u00bc 1 rad/sec. The simplified circuit is shown in frequency domain in Figure 4.10.3. The impedances of the components are as follows: Z1 H \u00bc j\u03c9L \u00bc j 1 1 \u00bc j \u03a9 \u00f02\u00de 180 4 Solutions of Problems: Sinusoidal Steady-State Analysis of Circuits. . . Z2 H \u00bc j\u03c9Leq \u00bc j 1 2 \u00bc j2 \u03a9 \u00f03\u00de Z2 \u03a9 \u00bc 2 \u03a9 \u00f04\u00de Therefore: Zab \u00bc j\u00fe j2\u00fe 2 \u00bc 2\u00fe j3\u00f0 \u00de \u03a9 Choice (2) is the answer. 4.12. The circuit of Figure 4.11.1 illustrates the primary circuit in frequency domain. The impedance of each inductor is as follows: Z1 H \u00bc j\u03c9L \u00bc j\u03c9 1 \u00bc j\u03c9 \u03a9 \u00f01\u00de To determine the inductance matrix of the circuit, we need to find a relation between the phasors of the primary and secondary voltages and currents of the circuit in the following matrix form: V0 1 V0 2 \u00bc j\u03c9 L11 L12 L21 L22 I01 I02 \u00f02\u00de As we know, the relation below exists between the primary and secondary voltages, currents, and the numbers of turns of a transformer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001767_9781118316887.ch8-Figure8.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001767_9781118316887.ch8-Figure8.12-1.png",
        "caption": "Figure 8.12-1: Two-pole, three-phase, permanent-magnet ac machine.",
        "texts": [
            "11-1 Express Vq r s and VJS for Lq ? Ld. [Vq r s = rsFqs + urLdFds + ur\\*; VI = rsFds-urLqFqs} 392 PERMANENT-MAGNET ac MACHINE SP8.11-2 A permanent-magnet ac machine is controlled to operate with Fds = 0. Lq = 1.5Ld. Express Te. [Te = f X'^s] SP8.11-3 The machine given in SP8.11-1 is controlled to operate in the constant-power mode. The torque is expressed by (8.11-9). Why? [Ir ds is not zero in the constant-power mode] MAGNET ac MACHINE A two-pole, three-phase, permanent-magnet ac machine is shown in Fig. 8.12-1. The stator windings are identical windings, displaced by |7r. The windings are sinusoidally distributed, each with Ns equivalent turns and resistance rs. Electromagnetic torque is produced by the interaction of the poles of the permanent-magnet rotor and the poles resulting from the rotating air-gap mmf established by currents flowing in the stator windings. The rotating air-gap mmf (mmfs) established by symmetrical three-phase stator windings carrying balanced three-phase currents is given by (4.4-18). Voltage Equations and Winding Inductances The voltage equations for the two-pole, three-phase, permanent-magnet ac machine shown in Fig. 8.12-1 may be expressed as 8.12-1) 8.12-2) 8.12-3) 8.12-4) 8.12-5) Vas Vbs 11\u2014 Ts^as ~ ^s^bs = r\u201e?' \u201e + + \u25a04- d\\as ~dT dXbS ~dT dXcs Jcs ' s\u00b0cs dt In matrix form, Vabcs = \u00cf\"s\u0302 -abcs H~ P^abcs For voltages, currents, and flux linkages, y^abes) [Jas Jbs Jcs\\ and 8.12. THREE-PHASE PERMANENT-MAGNET ac MACHINE 393 394 PERMANENT-MAGNET ac MACHINE r, = 0 0 0 0 rs 0 0 rs The flux linkage equations may be expressed as Aas \u2014 J-'asas \"as ' \u25a0*-Jasbs*'bs ' \u25a0LJascs<jcs + Aa (8.12-6) (8.12-7) Abs \u2014 l^bsas^as \" J^bsbs^bs ~T~ -^bscs^cs ' ^bsm (8",
            " Torque An expression for the electromagnetic torque may be obtained by using the second entry in Table 2.5-1. Since the magnetic system is assumed to be linear, the field and coenergy are equal: Lis + Lms - i f tr 2\u00b1Jrns 1 T 21Jms Lis + Lms - i f 2\u00b1J,ms -IT Lis + Lms J 8.12. THREE-PHASE PERMANENT-MAGNET ac MACHINE 395 Wc = -Lss(ils + i\\a + i2 cs) - -Lms(iasib3 + iasics + ibsics) + \\'miassm.6r + X'mibs sin(0r - \u00a77r) + \\'mics sin(0r + \u00a77r) + W} pm (8.12-13) where Wpm relates to the energy associated with the permanent magnet, which is constant for the machine shown in Fig. 8.12-1. Taking the partial derivative with respect to 6r yields P 1 1 P\\ (ias ~ -zibs ~ -^cs) cos(9r + -^-(ibs - ics) sin(9r (8.12-14) The above expression is positive for motor action. The torque and speed may be related as given by (8.4-3). Machine Equations in the Rotor Reference Frame Since there are three stator variables (/as, /&s, and / c s ) , we must use three substitute variables in the transformation of the stator variables to the rotor reference frame. In particular, where lqdOs \u2014 J : vs ia6cs ( V O s ) \u2014 [fqs fds J0s\\ (8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003933_amr.328-330.520-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003933_amr.328-330.520-Figure2-1.png",
        "caption": "Fig. 2 A nonstandard metal gear [11]",
        "texts": [
            " By contrast with crystalline powder, laser sintering of amorphous powder is achieved at substantially lower pulse energy for its low melting point. Theory and algorithms of neural network are applied for studying SLS process parameters to optimize SLS. Such process parameters as layer thickness, hatch spacing, laser power, scanning speed, work surroundings temperature, interval time, and scanning mode are obtained by adopting genetic algorithm based on neural network model [8-10] . Fig.1 is the principle of SLS process. Fig.2 is a nonstandard metal gear manufactured by SLS [11] . Selective Laser Melting (SLM). SLM was developed based on SLS. The same as SLS, SLM is only suitable for small parts with complex shape. Different from SLS, SLM can produce metal products with increased density significantly. To alleviate glomeration of molten metal, material parameter, process parameter and scanning mode should be controlled strictly. Hot isostatic pressing (HIP) is required to obtain high density parts, which often increases producing difficulty, time and cost"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002357_ecce.2013.6647276-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002357_ecce.2013.6647276-Figure10-1.png",
        "caption": "Fig. 10. Model 3 for the 2-pole/12-slot configuration.",
        "texts": [
            " The variation of the magnetic flux density depending on the magnetized rotational angle \u03b8m still remains in both models of the 4-pole/12-slot configuration. D. 2-pole/12-slot Configuration To reduce the influence of the magnetized rotational angle on the torque ripple variation, a 2-pole/12-slot configuration was investigated and is reported in this section. When models 1 and 2 were applied to the 2-pole motor, the phase current direction on one side shown in Fig. 6 was reversed. In addition to the aforementioned models 1 and 2, model 3 is discussed as shown in Fig. 10. This winding arrangement has previously been proposed by the authors with the intent of reducing interference and variation of the magnetic suspension force in a bearingless motor [7]. This winding arrangement contributes to reduction in the harmonic components of the stator\u2019s MMF distribution. Additional short-pitch sub-coils were wound around the stator teeth on both sides of the main coil. The number of turns of the main coil was decreased to accommodate the sub-coils, e.g., one main coil of the U-phase and one sub-coil of the V- or W-phase were wound together in one tooth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003750_amm.197.764-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003750_amm.197.764-Figure1-1.png",
        "caption": "Fig. 1: Dimensions of workpiece. Fig. 2: Dimensions in cross-section.",
        "texts": [
            " Zhao, Cho and Kim [7] developed a three-dimensional simulation technique of a time-dependent finite element analysis through Lagrangian formulation. The computational results presented good coincidences with the corresponding experiments of laser cladding process. In this study, we set up a three-dimensional finite element model to simulate the temperature variation during laser cladding. The powder feeding during cladding is realized by setting the related elements into \u201cdeath\u201d and \u201cbirth\u201d. This enhances the simulation accuracy. The dimensions of the workpiece are shown in Fig.1 and Fig.2. The material of substrate is 10 mild carbon steel (AISI1010). The clad layer is 1Cr18Ni9 stainless steel (AISI302). The thermal properties of materials are temperature-dependent [8]. The heat convection coefficient between workpiece and surrounding is 6Wm -2 K -1 . Half of the workpiece is calculated as the symmetry. The object is meshed with 8 nodes hexahedral isoparametric elements (Fig.3). The mesh around clad is refined to assure the analysis precision. 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 TTP, www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002643_amr.308-310.1714-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002643_amr.308-310.1714-Figure3-1.png",
        "caption": "Fig. 3: 1/4 FE model of the elastic shaft coupling Fig. 4: Stress caused by torque",
        "texts": [
            "The outer diameter and maximum inner diameter of the elastic shaft coupling are \u00d87mm and \u00d83mm according to the known parameters shown in Table 1. The axial distance between two groups of grooves and that between two pairs of grooves in one group are 2mm and 1mm respectively, and therefore c=(3+l)mm. Fig. 2 shows the maximum positive stress on the elastic shaft coupling as a function of the height and length of the thin wall according to Eq. 6. All kinds of height with corresponding length can meet strength condition. The height of 0.4mm and length of 2mm were primarily chosen. A 1/4 finite element (FE) model (Fig. 3) of the elastic shaft coupling with solid45 element type was founded in a finite element package ANSYS and the FE analysis was conducted. The maximum Von Mises stress caused by deviation of 0.1mm is about 361.226MPa (Fig. 6), which is far more than that by torque of 5N\u00b7mm (Fig. 4) and that by axial force of 5N (Fig. 5). The maximum stress caused by the combined load (Fig. 7) is about 366.132MPa, which is less than allowable stress and close to that by deviation and so the strength of the elastic shaft coupling is mainly determined by deviation amount"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001704_978-88-470-2056-6_17-Figure17.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001704_978-88-470-2056-6_17-Figure17.1-1.png",
        "caption": "Fig. 17.1. Deformation of an infinitesimal parallelepiped",
        "texts": [
            "13) He achieves this result both directly, by developing consistently the variation of dxdydz, or with the use of kinematic relations already obtained by Euler [104],10 that in an elementary parallelepiped with sides dx,dy,dz the variation of their size is expressed by the relation, correct up to infinitesimals of second order: dx ( 1+ \u2202\u03b4x \u2202x ) , dy ( 1+ \u2202\u03b4y \u2202y ) , dz ( 1+ \u2202\u03b4z \u2202 z ) , (17.14) that with a modern language, using the axial components \u03b5x,\u03b5y,\u03b5z of the tensor of deformation, and making reference to Fig. 17.1, can be rewritten as: dx(1+\u03b5x) , dy(1+\u03b5y) , dz(1+\u03b5z) . (17.15) 10 p. 286. Lagrange finally reaches the equilibrium equations of incompressible fluids, already obtained by Euler [103]11 and acknowledges that the multiplier \u03bb represents the scalar value of the pressure in the individual points of coordinates x,y and z: d\u03bb dx = \u0394X , d\u03bb dy = \u0394Y, d\u03bb dz = \u0394Z, (17.16) where \u0394 is the density of the fluid. After Lagrange, the application of the laws of virtual work to continuous bodies is still considering them as aggregates of molecules, and then as if it were a set of material points"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003653_iceee.2012.6421204-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003653_iceee.2012.6421204-Figure2-1.png",
        "caption": "Figure 2. The UAV and the frames",
        "texts": [],
        "surrounding_texts": [
            "The aerodynamic forces of lift (L) and drag (D) are de ned as [2], [4]: L = 1 2 SCL (VR) 2 (6) D = 1 2 SCD (VR) 2 (7) Where is the air\u00b4s density, VR is the relative speed of the wind, and S is the super cial area. The motors forces are de ned as: FMi = Ki!i (8) Where Ki is a constant that includes the physical aspects of the propellers (angle of attack, diameter, form, etc) and !i is the angular speed of every motor. Every motor generates a couple of drag that is opposite to the movement of rotation of the propeller and it is proportional to the angular speed of every motor and is de ned as: DMi = KDi!i (9) Where KDi is a constant that includes the physical aspects of the propellers."
        ]
    },
    {
        "image_filename": "designv11_100_0002970_amm.86.619-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002970_amm.86.619-Figure1-1.png",
        "caption": "Fig. 1 Three-dimensional housing model",
        "texts": [
            " On the theoretical foundation of solid element mechanics, this model is analyzed by ANSYS and assessed in housing rigidity. Weak points in the previous designs have been detected. To overcome them, feasible pertinent measures have been taken through analysis and test. Thus, an optimal design comes into shape. In static force analysis inside housing, how to select loading and regulation methods is of vital importance. Only a right selection can ensure a most proximity to simulated actual working conditions. A three-dimensional model is built (as shown in Fig. 1). Force analysis has two steps. Step 1: Cohesive joints are selected and their rigidity matrix is worked with ANSYS. Step 2: The matrix is introduced into MASTA to work out the results of forces on each cohesive joint from loads, as shown in Fig. 2 below. Housing is gridded into 8-joint-and-6- surface units with FEA. It follows 2 steps. Step 1: general gridding and calculation. Step 2: detailed gridding (as shown in Fig. 3) based on force situation. The stronger forces are, the closer grids are"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.17-1.png",
        "caption": "FIGURE 1.17. Illustration of a wheel attched to the spindle axle.",
        "texts": [
            "14. O set is the distance between the inner plane and the center plane of the rim. A rim may be designed with a negative, zero, or positive o set. A rim has a positive o set if the spider is outward from the center plane. Rim width DC rim WDC rim Rim diameter Rim width Rim diameter 7 \u00bd \u2013 JJ 15 55 5 \u2013 114.3 The ange shape code signi es the tire-side pro le of the rim and can be , , , , , , , , , and . Usually the pro le code follows the nominal rim width but di erent arrangements are also used. Figure 1.17 illustrates how a wheel is attached to the spindle axle. Example 37 Wire spoke wheel. A rim that uses wires to connect the center part to the exterior ange is called a wire spoke wheel, or simply a wire wheel. The wires are called spokes. This type of wheel is usually used on classic vehicles. The highpower cars do not use wire wheels because of safety. Figure 1.18 depicts two examples of wire spoke wheels. Example 38 Light alloy rim material. Metal is the main material for manufacturing rims, however, new composite materials are also used for rims"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.31-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.31-1.png",
        "caption": "Figure 12.31 Poinsot geometric interpretation of possible angular velocity and momentum.",
        "texts": [
            "427) states that the inner product of angular momentum BL = BI B G\u03c9B by itself is constant: BL \u00b7 BL = L2 = I 2 1 \u03c92 1 + I 2 2 \u03c92 2 + I 2 3 \u03c92 3 (12.475) which shows that the magnitude and direction of BL are both constant. The constant vector BL indicates an invariant plane in the B-frame. The projection of B G\u03c9B on the invariant direction of BL is constant. It follows that any increment \u03c9 between two arbitrary moments of time is perpendicular to BL: BL \u00b7 B G\u03c9B = 0 This equation defines the invariant plane. This geometric interpretation of the integrals of motion and the invariant plane is illustrated in Figure 12.31. Example 749 Duffing Equation Representation Let us take a time derivative of torque-free Euler equations (12.405)\u2013(12.407): \u03c9\u03081 = b1 (\u03c9\u03072\u03c93 + \u03c92\u03c9\u03073) (12.476) \u03c9\u03082 = b2 (\u03c9\u03073\u03c91 + \u03c93\u03c9\u03071) (12.477) \u03c9\u03073 = b3 (\u03c9\u03071\u03c92 + \u03c91\u03c9\u03072) (12.478) b1 = I2 \u2212 I3 I1 b2 = I3 \u2212 I1 I2 b3 = I1 \u2212 I2 I3 (12.479) Substitution of \u03c9\u03071, \u03c9\u03072, \u03c9\u03073 from (12.405)\u2013(12.407) provides \u03c9\u03081 = b1\u03c91 ( b2\u03c9 2 3 + b3\u03c9 2 2 ) (12.480) \u03c9\u03082 = b2\u03c92 ( b3\u03c9 2 1 + b1\u03c9 2 3 ) (12.481) \u03c9\u03073 = b3\u03c93 ( b1\u03c9 2 2 + b2\u03c9 2 1 ) (12.482) We can solve (12.474) and (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001918_9783527645442.ch14-Figure14.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001918_9783527645442.ch14-Figure14.6-1.png",
        "caption": "Figure 14.6 Structure of the cross-linked polyelectrolyte gel membrane and the reversible change of the chemical stimuli-responsive membrane in the absence (a) and presence (b) of cholesterol. (c) Faradaic impedance spectra obtained for the membrane-modified electrode in the absence (curves a,c) and presence (curves",
        "texts": [
            " Further advances in the design of signal-responsive materials and their immobilization on solid supports [88, 89] resulted in the new switchable electrodes reversibly activated and deactivated by chemical or biochemical signals [19, 50]. For example, a nanostructured membrane prepared from a polyelectrolyte, quarternized poly(2-vinylpyridine (qP2VP) cross-linked with 1,4-diiodobutane (DIB), was deposited on an ITO electrode surface, and the membrane pores were reversibly closed and opened upon reacting the membrane with cholesterol and washing it out, respectively [50] (Figure 14.6a,b). Cholesterol binding to the polymer chains through the formation of hydrogen bonds resulted in the swelling of the membrane and closing of the pores (Figure 14.6b), while washing of cholesterol from the membrane restored its initial state with the open pores (Figure 14.6a). The reversible opening/closing of the membrane pores was followed by measuring the electrode electron transfer resistance for a diffusional redox probe using Faradaic impedance spectroscopy (Figure 14.6c). The cyclic addition and removal of cholesterol to and from the membrane resulted in the modulated interfacial resistance controlled by chemical signals (Figure 14.6c, inset). pH-sensitive polymer brushes [88] tethered to electrode surfaces demonstrated pH-switchable interfacial behavior when a neutral state of the polyelectrolyte was impermeable for ionic redox species, keeping the electrode mute, while the ionized state of the polymer brush was permeable for ionic redox species of the opposite charge, allowing their access to the conducting support and activating the electrochemical process [19, 90, 91]. Protonation\u2013deprotonation of the polymer brush was controlled by the pH value, which could be changed in the bulk solution [19, 90, 91] or varied locally at the interface by means of electrochemistry (Section 14"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.10-1.png",
        "caption": "Figure 2.10. Representation of the",
        "texts": [
            " This justifies the approach of new variables more adequate comparative to the stator phase quantities: ia, ib, ic, va, vb, vc, and so on. This indispensable transformation is given by Park transformation. 2.1.3.2 Park Transformation. The transformation proposed by Park is based on the two-reaction theory originally developed by Blondel [3] and the further exposition of the concept by Doherty and Nickle [4]. The Park transformation, proposed initially in 1929 [5], consists in replacing the three stator windings a, b, and c by three fictitious rotor windings called d, q, 0 (rotor coordinates system) (Figure 2.10): The windings disposed in the d-axis and q-axis are rotating together with the rotor, and there are no mutual couplings between them. The 0-axis is independent of the d-axis and q-axis since zero-sequence currents flow through it only in unbalanced conditions. The main advantage of the Park transformation is that all windings from Figure 2.10 are stationary to each other, thus self- and mutual inductances are constant in time. This observation helps considering more simple equations using d, q, 0 components instead of phase quantities a, b, c. Considering the flux equations (2.21) of the rotor circuit windings cf \u00bc Lf if \u00fe LfDiD Lmd ia cos u \u00fe ib cos u 2p 3 \u00fe ic cos u \u00fe 2p 3 cD \u00bc LDf if \u00fe LDDiD LmD ia cos u \u00fe ib cos u 2p 3 \u00fe ic cos u \u00fe 2p 3 cQ \u00bc LQQiQ \u00fe LmQ ia sin u \u00fe ib sin u 2p 3 \u00fe ic sin u \u00fe 2p 3 it can be seen that the stator currents are similarly combined in each axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003509_rast.2013.6581261-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003509_rast.2013.6581261-Figure1-1.png",
        "caption": "Fig. 1 Defmition of the used frames",
        "texts": [],
        "surrounding_texts": [
            "The two main terms of the kinematical orientation of the sate II i te are: - The angular velocity of the BF related to the RF: ro BR = PBi + qBj + rBk (1) - The angular velocity of the RF related to IF: roRf = lU)R +lUjjR +lUkkR (2) The angular velocity of the RF related to IF ro Rf expressed in BF is: (3) where, the rotation matrix Ae will be further defined Finally we define the angular velocity of the BF related to the IF : rom =ro\ufffd+ro\ufffd \ufffd Because the RF are defined by unitary vectors: iR,jR,kR , their definition based on position vector r and velocity v: k r . vx r . . k R =- -;JR =-I --I;IR =hX R (5) r vxr Deriving in reference frame we obtain successive: diR \u2022 k dkR \u2022 \u2022 dt = co d R - co J R'\" dt = CO} R - co i h . (6) If we perform scalar product of each relation with unitary vectors i R' j R' k R ' found that: dkR \u2022 dkR \u2022 djR \u2022 co I = - dt h; co J = dt I R ; co k = - dt I R (7) Taking into account by definition relations (5) and the features of mixed product we obtain successively: co = 0 . lU = - h/r2 . co = 0 (8) I ' J ' k Obviously, for Keplerian movement, no acceleration outside from the orbit plan, so we can write: lURI = [0 lUI of (9) Taking into account the relation (4) we can write finally: [PB qB rB Y = [lUx lUy lUJ - A. [O lUj Of (10) This relationship is important because by measuring or calculating the angular velocity of the BF related to IF ro BI ' and using then (10) we can determine the angular velocity of the BF related with RF ro BR that can be used to control the vehicle attitude related to RF"
        ]
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure2-1.png",
        "caption": "Figure 2. Steel Press Center Structure.",
        "texts": [],
        "surrounding_texts": [
            "The big trend in the automotive industry is environment friendly vehicles developed from utilizing alternative energy technologies and fuel reduction through weight lightening technologies. Until now environment friendly vehicles that do not have mass production system have been focused on technology development. But vehicle weight reduction technology has been developed with the steady research and development because its fuel efficiency and impact.\nOn the other hand, away from commonly used steel and aluminum, high-strength lightweight engineering plastics industry has been developed because the effect is very good material.\nHowever, engineering plastics is significantly lower when compared to the value of young's modulus of steel and aluminum. So the stiffness for suspension components is considered to be an important factor.\nIn addition plastic material is changed depending on the heatresistant, fire-resistant characteristics property. So the current techniques are very difficult to preserve permanent properties as steel and aluminum. So if only engineering plastics is used as the suspension parts, problems of degrading the performance and quality will happen.\nIn the papers, the composite material suspension arm could be developed highlighting advantage of light weight of plastic and compensating for disadvantage of low strength of plastic. The concept of a composite material suspension arm is that with using a dual material of plastic and steel, plastic and steel is a technique for combining properties and large force is absorbed to steel and plastics compensates for steel.\nIn the papers, such as a Fig 1 is development of a composite material suspension arm that are based on upper arm of genesis suspension system. A composite material suspension arm is developed with the same strength and stiffness compared with forged steel upper arm. And durability verification is proved the part and vehicle durability test based on Hyundai Motors specifications.",
            "The basic structure of composites suspension arm fixed to mold from steel press and plastic is injected into mold.\nHowever it is easily separated if the coefficient of friction, surface and material properties in the structure are different from other structure.\nSteel press center structure is a strong combination because burring fix vertical direction as shown in Figure 3. Steel press center structure is a strong combination because flange is fixed horizontal direction as shown in Figure 3. So the structure of steel and plastic were fixed in any direction.\nIn general, ball joints that deliver the force from the tires are important parts with higher strength and durability. Also because ball stud often causes oscillation and rotation, ball joints bearings must be strong structure to withstand the steady force.\nBall stud and steel pipes are fixed to the mold and plastic is injected into the mold in order to increase the strength and stiffness for ball joint part such as a Figure 4.\nFor ball stud in order to smooth rotate and oscillate, plastic and bearings should be well fixed. However, because using the same materials of plastic and bearing, do not be combined with each other structure. So plastic and bearing slip will occur.\nTo prevent slip for vertical and horizontal rotation, upper and lower bearings surface were applied to the \u2018 \u2019 shape. And to combine ball stud, upper and lower bearings were designed as a locking structure as shown in Figure 5.\nSuspension bush is like the joints for human. And the role of the suspension bushings is absorbed the forces caused ball joint.\nBut because bushing outer pipe material is steel or aluminum, bushing and outer pipe occur to slip and each other is lack of pull-out strength.\nIf it is covered plastics after nurring a steel pipe, steel pipes and plastics have a strong combination and pull-out strength degradation does not occur, as shown in Figure 6. If it is designed in shown a Figure 6, current forged upper arm of production is equal to pull-out strength.",
            "As shown in Figure 7, Ball joint pipe and bushing pipe were welded to center steel press and in order to increase in the durability and strength of welds, welds part were covered with plastics.\nEngineering plastics is materials that improved which thermal properties and mechanical strength. But even if it has a heat resistant, chemical resistant and high strength, engineering plastic is absolutely necessary considering of the safety factor properties because it cannot have permanent properties like steel and aluminum.\nEngineering plastics have advantages that various properties can be implemented depending on the glass fiber content. In this paper, department of Plastic is PA6 and properties of glass fiber is 60 percent. The higher glass fiber content, highstrength performance can be implemented, but current technology has the highest content of a GF 60%.\nEngineering plastics become significantly different properties depending on in made of any direction glass fibers. Material properties become significantly different depending on perpendicular direction and parallel direction.\nEngineering plastic properties as a general expression are shown properties of glass fiber made in a parallel direction. But because the strength in perpendicular direction was approximately 30% weak compared to the parallel direction, properties considering of perpendicular direction was absolutely necessary."
        ]
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.44-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.44-1.png",
        "caption": "Figure 12.44 A clock with two unbalanced rotating arms in (a) and balanced rotating arms in (b).",
        "texts": [
            "665) \u23a1 \u23a3 QX QY QZ \u23a4 \u23a6 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2 \u23a3 l 2 FZ sin \u03b8 \u2212 l 2 FZ cos \u03b8 Iz\u03b8\u0308 + l 2 FY cos \u03b8 \u2212 l 2 FX sin \u03b8 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5 \u23a6 (12.666) Let us substitute the force components from (12.664) to determine the components of the driving moment 0M0: \u23a1 \u23a3 QX QY QZ \u23a4 \u23a6 = \u23a1 \u23a2\u23a2\u23a2 \u23a3 0 0 ( Iz + m1l 2 4 ) \u03b8\u0308 + 1 2 mgl cos \u03b8 \u23a4 \u23a5\u23a5\u23a5 \u23a6 (12.667) Therefore, Q0 = QZ is the required torque to turn the arm about the Z-axis: Q0 = (Iz + 1 4m1l 2 ) \u03b8\u0308 + 1 2 mgl cos \u03b8 (12.668) Let us consider the motion of the long arm of a clock, as shown in Figure 12.44(a). The arm is supposed to turn with a constant angular velocity \u03c9: \u03c9 = \u03b8\u0307 = 2\u03c0 60 rad/min = 6 deg/min (12.669) Having zero angular acceleration means the motor of the clock should be able to apply a variable torque Q0, Q0 (\u03b8) = mga cos \u03b8 (12.670) where \u03b8 is measured from the horizontal axis, as shown in Figure 12.43(a). The torque Q0 is positive for 90 deg < \u03b8 < \u221290 deg or 0 < t < 30 min and negative for \u221290 deg < \u03b8 < \u2212270 deg or 30 min < t < 60 min. Besides the actuator torque Q0 and gravitation force mg, there are also a constant frictional torque Qf and a resistive viscous torque Qc = c\u03b8\u0307 that apply on the arm. So, a better model of the dynamics is Q0 = (Iz + m1a 2) \u03b8\u0308 + mga cos \u03b8 + Qf + c\u03b8\u0307 (12.671) which shows that we still need a fluctuating torque to have a constant angular velocity: Q0 (\u03b8) = mga cos \u03b8 + Qf + c\u03c9 (12.672) The motors of stationary clocks are supposed to apply a constant and uniform torque. The best way to reduce the effect of the variable term of Equation (12.672) is to shorten a. Having a counterbalance and designing the hands of the clock with a = 0 as are shown in Figure 12.44(b) can solve the problem. Example 759 Dynamics of a Four-Bar Linkage Linkages are connected multibodies that make a closed loop. The first and the last links of a linkage are usually connected to the ground link. Figure 12.45(a) illustrates a closed-loop four-bar linkage and Figure 12.45(b) illustrates the free-body diagrams of its links. The positions of the mass centers and the position vectors of 0ni and 0mi for each link are known. The Newton\u2013Euler equations for link (i) are 0Fi\u22121 \u2212 0Fi + mig J\u0302 = mi 0ai (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002429_amm.332.145-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002429_amm.332.145-Figure1-1.png",
        "caption": "Figure 1 The Robot Model developed in Solidworks",
        "texts": [
            " The image processing is developed by a laptop Toshiba Satellite A 300-1JH which has the subsequent jobs: web server, robot control and video server. 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-09/07/15,02:43:32) In order to determine the center of gravity of the proposed solution of robot assembling, the authors developed a model in Solidworks which is presented in Fig. 1. The model has the gripper without the mobile phone HTC. The problem to solve was in connection with the robot stability, when the mobile phone is placed in the robot gripper. The present work was developed as a helpful tool for home surveillance. The owner of the house can leave the robot home alone. The robot will inspect all the spaces in accordance with the owner requests. A special condition of this project was to use the Matlab environment for robot controlling, path planning, spherical bodies recognition and human faces detection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure14-1.png",
        "caption": "Figure 14 \u2212 Static tests",
        "texts": [
            " 12c) so that the moment of inertia of the castor, with respect to the steering axis, does not change with the fork configuration; its value was identified as 0.37kgm2 by means of a free oscillation test. Castor lateral stiffness at wheel spindle level The fork length difference, between the two extreme configurations, is equal to 150 mm and implies a significant lateral stiffness difference, that was evaluated by means of static tests. The castor, constrained to the rig frame with locked steering axis, was loaded in correspondence of the wheel axis (Fig.14); the displacements of points A (fork plate) and B (wheel axis) were detected for several load values and reported in figure 15. From load-displacement diagram relative to point B the lateral stiffness of about 9.0E4N/m can be deduced. This stiffness is evaluated at the wheel spindle level and therefore does not take into account the wheel flexibility but includes the flexibility of joint connecting the castor to the rig. 5 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002833_imece2013-62921-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002833_imece2013-62921-Figure3-1.png",
        "caption": "Figure 3: The connection area, center of area of the pins (O) and the shear forces.",
        "texts": [
            " The shear force applied to each pin ( ) is composed from two components: the resulting shear force created by uniform shear load (primary load ( \u2032 ) distribution on pins which is calculated for each one from equation (1), and the resulting force from the moment (M) (secondary shear force) which is proportional to the distance between the center of each pin and the central point of pins ( ). In other words, the closer a pin is from central point, the smaller is the secondary force on it. The secondary forces ( \" ) are calculated from Equation (2). It is noted that the assumption of the uniform distribution of the shear load V on pins requires members\u2019 rigidity, no existence of clearance and the similarity in pins` material and dimensions. (1) n V Fi  (2) 2 i n i i i r Mr F   (3) O xPM . Figure 3 shows the connection area where the center point of each pin is placed. The mathematical model of the optimization problem is as follows: Input Parameters Finite element modeling Finite element solution Yes No Convergence criteria check GA creates new generation Optimal design Input parameters Optimization using GA Optimal design Finite element modeling Finite element validation Mathematical modeling 2 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 10/21/2017 Terms of Use: http://www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003093_amm.204-208.4518-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003093_amm.204-208.4518-Figure1-1.png",
        "caption": "Fig 1 Coordination systems",
        "texts": [
            " Mathematic modal of the platform and a special program was created with M-language for the time-domain dynamic analysis. PID control strategy, kalman filtering theory and optimal thrust allocation method were used in the analysis. Simulation result indicated that the DPS of this platform is safe and efficient. Three coordination systems need to be used for describing the motions of the platform: global coordination system EEOYX , vessel coordination system BBOYX and parallel coordination system XOY , see Fig1. Platform in ocean have 6 motion freedom degrees. In the DPS research, just sway, surge and yaw need to be taken into consideration. For avoiding the power waste and reducing the wear of thrusters ,just low frequency motion need to be paid attention to. According to the maneuverable theory, nonlinear maneuverable equation of platform motion can be obtained. In the model of platform motion, it is supposed that coupling just existed in sway and yaw, and surge has no coupling with sway and yaw. On such points, simplified motion equation can be described as: ( ) ( ) ( ) x Xwa Xwi XC XT u y Ywa Ywi YC YT v r z z wa wi C T r v m m u F F F F X u m m v F F F F Y v Y r I J r N N N N N r N v  + = + + + +  + = + + + + +  + = + + + + + (1) All rights reserved"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.40-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.40-1.png",
        "caption": "Fig. 3.40",
        "texts": [
            " Without going into the details of the derivation it shall be noted that for each reference point there exists a specific coordinate system with three orthogonal axes 1, 2 and 3 for which all products of inertia are zero. These axes are referred to as principal axes (of inertia). The respective moments of inertia are extremal values and called principal moments of inertia. In the principal axes system the inertia tensor takes the simple form \u0398(A) = \u23a1\u23a2\u23a2\u23a3 \u03981 0 0 0 \u03982 0 0 0 \u03983 \u23a4\u23a5\u23a5\u23a6 (3.55) where \u03981, \u03982 and \u03983 are the principal moments of inertia. In the case of a homogeneous symmetric body, the symmetry axes are principal axes. Fig. 3.40a shows as an example the principal axes of a cuboid with respect to its center of mass. For axisymmetric bodies, the symmetry axis and any orthogonal axis are principal axes (Fig. 3.40b). In the special case of a homogeneous thin plate, the mass of an element is given by dm = \u03c1 t dA, see Fig. 3.40c. Since z is small compared with x and y, it can be neglected in the integration (z \u2248 0). It then follows from (3.53) that the moment of inertia with respect to the x-axis can be written as \u0398x = \u03c1 t \u222b y2 dA = \u03c1 t Ix . Here, Ix is the area moment of inertia with respect to the x-axis (cf. Volume 2). Applying this procedure for all components of the inertia tensor, we obtain \u0398x = \u03c1 t Ix, \u0398y = \u03c1 t Iy, \u0398xy = \u03c1 t Ixy, \u0398z = \u03c1 t (Ix + Iy) = \u03c1 t Ip, \u0398xz = \u0398yz = 0 . (3.56) Thus, in this case a direct relation exists between the mass moments of inertia and the area moments of inertia"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.67-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.67-1.png",
        "caption": "Fig. 2.67 4RPaPPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology R\\Pa\\kP\\\\Pa",
        "texts": [
            " 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No. 58 T ab le 2. 4 L im b to po lo gy an d co nn ec ti ng co nd it io ns of th e fu ll y- pa ra ll el so lu ti on s w it h no id le m ob il it ie s pr es en te d in F ig s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001810_9781118120118.ch2-Figure2.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001810_9781118120118.ch2-Figure2.1-1.png",
        "caption": "Figure 2.1. Phase diagram of the ADE model [15] . The horizontal axis is the volume - to - area ratio expressed by the reduced volume v (scaled to that of a sphere). The vertical axis shows the effective differential area expressed by \u0394 a 0 (scaled to that of a sphere). Along the vertical axis, vesicles can change their shape by modifying the surface area difference between monolayers. Reprinted from Current Opinion in Colloid & Interface Science, 5, Hans - G \u00fc nther D \u00f6 bereiner, Properties of giant vesicles, 256 \u2013 263, Copyright (2000), with permission from Elsevier.",
        "texts": [
            " \u03b1 has been estimated to be of the order of one for all phospholipids. Equation 2.2 represents the area difference elasticity ( ADE ) model [14] . Again, the minimization of Equation 2.2 , at given parameters of volume - to - area ratio and preferred differential area, leads to the same diversity of vesicular shapes, but new pear, budded, starfi sh vesicles or vesicles with holes are now also predicted. Vesicle shape regions of minimal energy can be represented into a diagram, which contains different vesicle classes. An ADE shape diagram is reproduced in Figure 2.1 from Reference 15 . Shape changes of vesicles triggered by an asymmetric distribution of lipids constitute an experimental evidence of the bilayer structure included on the ADE model (see next section). This is equivalent to maintain constant the vesicle volume and to modify the preferred area difference along the vertical axis of the shape diagram. Lipid fl ip - fl op represents a very simple mechanism to vary the number difference in the ADE contribution (Eq. 2.2 ). Following the shape diagram shown in Figure 2.1 , it is evident that one could cross the different shape regions by modifying the spontaneous curvature and the preferred differential area of vesicles. Any asymmetry across the membrane will lead to a change in the effective spontaneous curvature. A large variety of mechanisms for monolayer surface area asymmetry generation can be found in the literature [16 \u2013 18] . Among all of them, an asymmetric distribution of lipids between leafl ets has been extensively explored in cells as well as in model vesicular systems",
            " On the other hand, when PG inward transport is induced, inverting the pH gradient, noninvaginated LUVs reverse to invaginated vesicular structures. All the experiments described above inspired the technique of shape changes performed on giant vesicles to quantify the transbilayer diffusion of unlabeled lipids presented in the next section. Shape of GUVs can be easily modifi ed by varying experimentally the parameters that describe their shape: the preferred area difference between monolayers (asymmetry) and the vesicular surface area/volume ratio (see Fig. 2.1 ). GUVs are very sensitive to the relative area between monolayers and the minimum asymmetry needed to induce a shape change is of the order of 0.1% of the total area of the vesicle [23, 24] . GUVs are consequently, in comparison with other membrane models, excellent objects to detect lipid asymmetry. They are visible under optical microscopy, which facilitates their observation and manipulation, and their internal volume can be modifi ed. This is important since the reduced volume must be lower than 1 (which corresponds to a sphere) in order to have a vesicular surface excess for allowing shape changes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure10-1.png",
        "caption": "Figure 10 \u2212 Castor handlebar",
        "texts": [
            " The test rig allows hence to set: a) the belt speed; b) the rake angle; c) the vertical load; d) the fork suspension extension; e) the steer damping. Since the force exerted by the steering damper depends almost linearly on the steer rotating speed, the damping value for each of the first nine damper positions (clicks) was estimated. Starting from the damper characteristic curves and taking into account the distance of the damper axis from the steering axis (124mm), the rotational damping coefficients \u03c3, for each damper click, was deduced and reported in table II. Figure 10 shows a photo of the castor handlebar equipped with the steer damper and the angular sensor. Castor lateral stiffness at wheel spindle level approximately depends on the cubed fork length and it is therefore strongly influenced by the fork deflection. In turn, fork deflection depends on the vertical load acting on the castor. A relationship between the vertical load and lateral stiffness is quite complicated to derive for the presence of frictional force arising between fork tubes and sliders. In figure 11, the loadtravel experimental diagram for one telescopic suspension, in vertical axis configuration, is reported: it may be noted that there are different deflection values corresponding to a given vertical load"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.35-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.35-1.png",
        "caption": "Figure 12.35 A top.",
        "texts": [
            " So, following, the same procedure, we find their solutions as \u03c91 (\u03b1) = \u03c910 cos \u03b1 + \u03c920 sin \u03b1 + \u222b \u03b1 0 ( M1 I1\u03c93 cos (\u03c4 \u2212 \u03b1) + M2 I1\u03c93 sin (\u03c4 \u2212 \u03b1) ) d\u03c4 (12.593) \u03c92 (\u03b1) = \u03c920 cos \u03b1 \u2212 \u03c920 sin \u03b1 + \u222b \u03b1 0 ( M1 I1\u03c93 sin (\u03c4 \u2212 \u03b1) + M2 I1\u03c93 cos (\u03c4 \u2212 \u03b1) ) d\u03c4 (12.594) Example 755 Top Dynamics A top is an axisymmetric rigid body that is rotating about a fixed point in a constant gravitational field. The fixed point is on the axis of symmetry at a distance l from the body mass center C. We attach a body coordinate frame B and a global frame G at the fixed point O. Figure 12.35 illustrates a top with mass moment [I ]: BI = \u23a1 \u23a3 I1 0 0 0 I1 0 0 0 I3 \u23a4 \u23a6 (12.595) Let us choose Euler angles \u03d5, \u03b8 , \u03c8 as the generalized coordinates to describe the motion of B in G. The rotation of the top about its axis of symmetry is the spin \u03c8 . The angle between the axis of symmetry and the Z-axis is the nutation \u03b8 , and the rotation of the axis of symmetry about the Z-axis is the precession \u03d5: BRG = \u23a1 \u23a3 c\u03d5c\u03c8 \u2212 c\u03b8s\u03d5s\u03c8 c\u03c8s\u03d5 + c\u03b8c\u03d5s\u03c8 s\u03b8s\u03c8 \u2212c\u03d5s\u03c8 \u2212 c\u03b8c\u03c8s\u03d5 \u2212s\u03d5s\u03c8 + c\u03b8c\u03d5c\u03c8 s\u03b8c\u03c8 s\u03b8s\u03d5 \u2212c\u03d5s\u03b8 c\u03b8 \u23a4 \u23a6 (12",
            " Figure 12.38 shows the path of the z-axis when \u03d5\u0307 (t) is changing between negative and positive values. Figure 12.39 depicts the situation in which \u03d5\u0307 (t) is changing between zero and a positive maximum. The periodic dance motion of the top with \u03b8 (t) which is superimposed on the precessional motion is called nutation. Example 756 Special Cases of Top Dynamics Some special cases of top dynamics have simpler equations of motion with simpler interpretations. Consider the top kinematics shown in Figure 12.35 and its dynamic equations of motion (12.603)\u2013(12.605): 1. \u03c93 = 0. This case is equivalent to a planar pendulum in which \u03b8 is the only time-dependent variable. The equations of motion reduce to only one equation: I1\u03b8\u0308 \u2212 mgl sin \u03b8 = 0 (12.632) The planar pendulum case has energy conservation as the only integral of motion: I1\u03b8\u0307 2 + 2mgl cos \u03b8 = 2E (12.633) 2. \u03b8 = \u03b80 = const. This case is dependent on constant precession angular velocity. Equation (12.608) yields \u03d5\u0307 = const and Equation (12.605) leads to \u03c8\u0307 = const"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002965_amr.328-330.743-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002965_amr.328-330.743-Figure2-1.png",
        "caption": "Fig. 2 Box-taking Mechanism",
        "texts": [
            " Among the mechanisms of realizing the trajectory, trajectory of four-bar can only realize the position of limited points accurately. Five-bar mechanism has two degrees of freedom, it can realize the complicated trajectory. Five-bar with a cam or the hybrid driven five-bar can realize the accurate trajectory, but it can not realize the trajectory with demand of velocity [5]. In this paper, design a five-bar with double cams to realize the accurate trajectory. The mechanism utilizes double cam groove driven five-bar mechanism. As shown in Fig.2. The Mechanism has four suction heads, only analyze one of them, the rest of them have same trajectory. Working principle: Cam plate 3 fixed in pedestal 2, connect Spindle 1with turntable 7. When spindle 1 revolves with turntable 7, five-bar mechanism rotates with turntable 7 together. Five-bar mechanism has two degrees of freedom and needs two input power to form precise trajectory. respectively through the linkages, and then drive two input power linkages of five-bar mechanism, making C end point in five-bar mechanism execute the designated action in designated position "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.46-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.46-1.png",
        "caption": "Fig. 3.46",
        "texts": [
            "61) we obtain Mx = \u0398xz \u03c9\u0307 \u2192 \u2212 b By = \u2212 \u03c9\u0307mcb 4 , (c) My = \u0398xz \u03c92 \u2192 b Bx = \u2212 \u03c92mcb 4 , (d) Mz = \u0398z \u03c9\u0307 \u2192 M0 = \u03c9\u0307mc2 6 . (e) The last equation represents the equation of motion. Assuming the initial condition \u03c9(0) = 0 leads to 3.4 Kinetics of a Rigid Body in Three Dimensional Motion 197 \u03c9\u0307 = 6 M0 mc2 \u2192 \u03c9 = 6 M0 mc2 t . Insertion into (a) - (d) finally yields Ax = \u2212 \u03c92mc 12 , Ay = M0 2 c , Bx = \u2212 \u03c92mc 4 , By = 3 M0 2 c . E3.23Example 3.23 A parasitic mass m0 is attached to the wheel of a car with rotation axis z as shown in Fig. 3.46. Determine the masses m1 and m2 that must be attached at locations 1 and 2 in order to balance the wheel. Solution The wheel is balanced if the center of mass is located on the rotation axis and if all the products of inertia vanish. Thus, with the notation according to Fig. 3.46 the following conditions must be fulfilled: m0 r0 + m2 r2 = m1 r1 , \u0398zy = \u2212m0 r0 e0 + m1 r1 e1 + m2 r2 e2 = 0 . Solving yields the masses m1 = m0 r0 r1 e0 + e2 e1 + e2 , m2 = m0 r0 r2 e0 \u2212 e1 e1 + e2 . A rigid body rotating arbitrarily about its center of mass C or a fixed in space point A (e.g. a pivot) is referred to as gyroscope. We call a gyroscope torque-free if the moment of the external forces about the reference point C or A vanishes. This condition is fulfilled e.g. for the rotation of a satellite"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002978_detc2011-48760-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002978_detc2011-48760-Figure2-1.png",
        "caption": "Figure 2. Tooth surface of complementary crown gear.",
        "texts": [
            " Therefore, the normal module is represented by g gm p pm n z R z R M 00 sin2sin2   (6) The purpose of this study is to manufacture the large-sized straight bevel gears with equi-depth on multi-tasking machine in the previous section. Therefore, the tooth surfaces of the pinion and gear can be modified arbitrarily. The tooth surface of the complementary crown gear can be also designed arbitrarily using this advantage. The number of teeth of the complementary crown gear is represented by 00 sinsin g g p p c zz z   (7) Figure 2 shows the tooth surface of the complementary crown gear. O-xyz is the coordinate system fixed to the crown gear and z axis is the crown gear axis. Point P is a basic point at which tooth surfaces mesh with each other and is defined in the Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/22/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use 3 Copyright \u00a9 2011 by ASME center of tooth surface. The circular arcs with large radii of curvatures are defined both in xz and xy planes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.14-1.png",
        "caption": "Fig. 3.14",
        "texts": [
            " The point of intersection of these straight lines is point \u03a0 (Fig. 3.13b). From the figure we obtain with the theorem of intersecting lines rA r1 \u03c91 = 2 r3 \u2212 rA r2 \u03c92 \u2192 rA = 2 r3 r1 \u03c91 r1 \u03c91 + r2 \u03c92 . The angular velocity \u03c93 of the disk follows from (3.9): vA = rA \u03c93 \u2192 \u03c93 = vA rA = r1 \u03c91 + r2 \u03c92 2 r3 . The velocity of point C is given by vC = (rA \u2212 r3)\u03c93 = 1 2 (r1 \u03c91 \u2212 r2 \u03c92) . In the special case of r1 \u03c91 = r2 \u03c92 the motion of the disk is a pure rotation (vC = 0). E3.4 Example 3.4 Link 1 of the mechanism in Fig. 3.14a rotates with angular velocity \u03c91. Find the velocities of points A and B and the angular velocities of links 2 and 3 at the instant shown. Solution Links 1 and 3 rotate about the points C and D, respectively (Fig. 3.14b). Thus, these points are the centers \u03a01 and \u03a03 of zero velocity of the respective links. Therefore, the directions of the velocities of A and B are known. The points A and B are also points of link 2 . Its center \u03a02 of zero velocity is given by the point of intersection of the two straight lines that are perpendicular to vA and vB. From Fig. 3.14b we obtain \u03a01 : vA = \u221a 2 a\u03c91, \u03a02 : vA = 2 \u221a 2 a\u03c92 \u2192 \u03c92 = \u03c91 2 , vB = 2 a\u03c92 \u2192 vB = a\u03c91, \u03a03 : vB = a\u03c93 \u2192 \u03c93 = \u03c91 . 3.2 Kinetics of the Rotation about a Fixed Axis 147 In Section 3.1 we studied the motion of a rigid body without referring to forces as a cause or as a result of the motion. Now we will investigate motions under the influence of forces. In this section we restrict ourselves to the rotation about a fixed axis. As the body in Fig. 3.15 rotates about the fixed axis a-a, each point of the body undergoes a circular motion"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002797_amr.468-471.2141-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002797_amr.468-471.2141-Figure8-1.png",
        "caption": "Fig. 8 The rolling of the cross-rail in the up-down direction",
        "texts": [],
        "surrounding_texts": [
            "Considering all the directions of the excitation, the simulated and experimental results and the error between the simulated and experimental results with joints have been provided in the Table.2 within the frequency of 100 Hz. It is found that the error between the simulated and experimental results of the whole assembled structure is less than 8.8%, while the error of the contact model often used in the existing literatures is one times bigger than the model.which indicates the proposed method in this paper is effictive. Table 2 Modal frequency and errors of the assembled structure Simulated result without joints [Hz] Simulated result with joints [Hz] Experimental result [Hz] Error [%] 23.318 17.689 19.388 -8. 763 70.602 50.418 46.409 8.638 90.211 84.851 81.439 4.190"
        ]
    },
    {
        "image_filename": "designv11_100_0002731_20110828-6-it-1002.03717-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002731_20110828-6-it-1002.03717-Figure1-1.png",
        "caption": "Fig. 1. 3D cutaway drawing",
        "texts": [
            " This estimate, coupled with a parameter estimation method, can compensate for variations in certain parameters of the LuGre model. The paper has been arranged in the following manner; actuator and friction modeling has been presented in section 2. The sliding mode observer and output injection have been explained in section 3. Simulation results have been presented in section 4, and experimental results have been presented in section 5. Some conclusions have been discussed in section 6. In modern diesel engines, the swirl actuator is integrated in the air inlet manifold. As seen in figure 1, the actuator consists of a DC motor, a return spring and gearing. The significant number of mechanical parts result in friction and stick-slip, spring nonlinearities, as can be seen in the 978-3-902661-93-7/11/$20.00 \u00a9 2011 IFAC 7480 10.3182/20110828-6-IT-1002.03717 static characteristic curve of the actuator in figure 2. The modeling hence depends upon estimation for parameter identification. A detailed and physically motivated simulation model has been defined in Ahmed et al. [2010a], in which motorspring systems have been characterized by nonlinearities due to the presence of spring pre-compression and significant friction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003011_amm.186.169-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003011_amm.186.169-Figure3-1.png",
        "caption": "Fig. 3 Schematic diagram of electric arc spray metallization",
        "texts": [
            " The cathode is made of tungsten alloyed with a small quantity, less than 1%, of thorium, having the role of unifying the electrode tip, which leads to pilot arc and plasma jet establishment. The anode is made of copper and is cooled with water in order to reduce heat strain and wear. The most used gas to obtain plasma jet is Ar. The schematic diagram of APS is shown in Fig. 2. APS layer properties are presented in Table 2. Tabel 2 APS layer properties Properties Value Surface adhesion strength for ceramic layers 25...30 MPa Surface adhesion strength for metal layers: Mo, NiAl, NiCrAl It can reach 70 MPa Schematic diagram of arc metal spraying is shown in Fig. 3 [2]. The two wires-electrode, 1, move forward through a push-pull type system. The electrode wires are elastic deformed and they led to approach each other to make a short circuit and so to form an arc between two stripes of copper, 2. The compressed-air which gets out through spray nozzle, 3, sprays the melted material from the tip of the electrode wires, 4, and together with the spray mist they form the spray jet, 5, which is normal oriented on the piece surface, 6. The layer properties obtained by metal arc spraying are presented in Table 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002378_s11232-013-0053-x-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002378_s11232-013-0053-x-Figure1-1.png",
        "caption": "Fig. 1. The schematic structure of a composite particle propagating in a constant electric field E: the internal particle with the charge e2 has the mass m2. The envelope is neutral, e1 = 0, and has the mass m1. We let V1 and V2 denote the respective velocities of the envelope and the internal particle.",
        "texts": [
            " We show that the billiard formalism allows finding the main features of the motion of such a particle and results in several interesting problems. Using the generalized trick of straightening trajectories proposed here, we show that also in this case, the motion of a charged composite particle reduces to that of a pointlike particle in an unbounded medium but in a potential energy field of special form. We consider the motion of a composite particle with one internal degree of freedom in a constant electric field E. Let the particle envelope be a cylinder of length L with closed tops having the mass m1 (see Fig. 1). An internal particle of mass m2 propagates freely inside the envelope. We first assume that only the internal particle has the charge e2 and the envelope is neutral or noncharged (e1 = 0). We describe general laws of motion of such a particle using the billiard approach [2]. We begin with equations of motion for the envelope and internal particle, m1x\u03081 = 0, m2x\u03082 = Ee2, (1) where x1 and x2 are the respective coordinates of the envelope center of mass and the internal pointlike particle. We then pass to the new coordinates x\u03031 = \u221a m1x1 and x\u03032 = \u221a m2x2, which we call the billiard coordinates in what follows"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001944_9781118354162.ch13-Figure13.38-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001944_9781118354162.ch13-Figure13.38-1.png",
        "caption": "Figure 13.38 A nitrogen oxide microsensor. The surface is coated with Nafion and WPI membranes. Adapted with permission from [89]. Copyright 2002 Elsevier.",
        "texts": [
            " Being a cation exchanger, Nafion prevents the access of the anionic species to the electrode but is permeable to the neutral nitric oxide molecule. However, Nafion is not effective in eliminating interference from cationic species (such as dopamine and serotonin) or neutral compounds (for example, hydrogen peroxide and acetaminophen). A second, hydrophobic membrane prevents the access of such species to the electrode surface but is not an obstacle to nitric oxide, which is a hydrophobic compound. Figure 13.38 displays an amperometric nitric oxide sensor designed according to the above principles. The working electrode consists of a carbon microfiber electrode (7mm diameter) with a conical active tip coated with Nafion and a hydrophobic WPI membrane. Integrated with this is an Ag/AgCl reference electrode coated as a thin layer over the carbon fiber holder. This sensor has a linear response to NO in a stirred buffer solution over the concentration range from 10 to 1000 nM. It is suitable for in vivo and real-time detection of nitric oxide in single cells or microsamples [89]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002691_imece2012-87321-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002691_imece2012-87321-Figure2-1.png",
        "caption": "Figure 2. INTERFACE POINTS FOR THE FRF CALCULATIONS",
        "texts": [
            " By comparing the FRFs of different systems it is possible to compare their dynamic behavior. The choice of the excitation and response points defines the investigated transfer path. The FRFs are derived from the multibody model by performing a linearization. With regard to reliability the main excitation source is the connection between the rotor and the main shaft. Three main loading degrees of freedom (DOFs): radial, thrust and bending are investigated. The input for each of these FRFS are the forces at the rotor main shaft coupling. Figure 2 shows the rotor-main shaft coupling and the corresponding forces. The main response DOFs are the planet carrier displacements at the rear planet carrier (PLC) bearing location (PLC-B) and the gearbox bushing displacements. Figure 1 shows the different bearings in the gearbox, whereas Figure 2 shows the gearbox bushings. The planet carrier motion is investigated, since this parameter is significantly affected by the non-symmetric loading of the gearbox. Figure 3 illustrates this. Due to the planet carrier displacements the loading conditions in the planet-ring gear meshes can become unfavorable. This can potentially result in non-symmetric planet bearing loading and corresponding overloading of one of the planet bearings [2]. The planet-ring gear meshes displacements are not used in the FRF analysis in this paper since we want to make abstraction of the influence of the gear meshing stiffness on the behavior of the gearbox suspension"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.7-1.png",
        "caption": "Fig. 2.7 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRPRR (a) and 4PRRRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R\\P\\kR\\R (a) and P\\R||R\\R||R (b)",
        "texts": [
            " 2.4a) R||R\\R||R||R (Fig. 2.1d) The two first revolute joints of the four limbs have parallel axes. 6. 4PRRRR (Fig. 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7. 4RRRPR (Fig. 2.5a) R||R\\R\\P\\kR (Fig. 2.1f) Idem No. 5 8. 4RRPRR (Fig. 2.5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10. 4PRRRR (Fig. 2.6b) P||R\\R||R\\R (Fig. 2.1i) Idem No. 9 11. 4RRPRR (Fig. 2.7a) R\\R\\P\\kR\\R (Fig. 2.1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2.8a) P||R||R||R||R\\R (Fig. 2.1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig. 2.9a) R||R||P||R\\R (Fig. 2.1n) Idem No. 13 16. 4RRRPR (Fig. 2.9b) R||R||R||P\\R (Fig. 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No. 13 18. 4RRPRR (Fig. 2.10b) R||R||P||R\\R (Fig. 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003778_978-3-642-17234-2_4-Figure4.19-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003778_978-3-642-17234-2_4-Figure4.19-1.png",
        "caption": "Fig. 4.19 A homogeneous cylinder rolling without sliding on an inclined plane.",
        "texts": [
            "55)1,2, the result is substituted into (4.6.55)3 and then one eliminates u between the obtained equation and (4.6.55)4. Consequently, _w sin2 h \u00bc C1 \u00fe C2 cos h; where C1 and C2 are two constants. Since we have obtained three first integrals, our problem has been reduced to quadratures. If the axis of rotation of the rolling sphere does not change its direction, then the constraint (4.6.53) becomes holonomic. This is the case, for instance, of a homogeneous cylinder, rolling without sliding on an inclined plane (Fig. 4.19). At the point of contact between the plane and the circle of transversal section with its centre in G, we have R = -Rk, such that: vG \u00bc Rx k; jvGj \u00bc _s \u00bc Rjx kj \u00bc Rx \u00bc R _h: If we choose s\u00f00\u00de \u00bc 0; h\u00f00\u00de \u00bc 0; we find s = Rh. A survey of the forces acting on the cylinder (see Fig. 4.19) leads to the Lagrangian: L \u00bc 1 2 \u00f0IG \u00feMR2\u00de _h2 \u00fe RMgh sin a; where IG is the moment of inertia of the cylinder relative to its axis. Then the Lagrange equations (2.5.17) yield: \u20ach 2g 3R sin a \u00bc 0: This equation, together with the initial conditions, determine uniquely the motion of the cylinder. Let us consider a system of charged particles (electrons), performing a finite motion in a central electric field, created by a fixed point charge. We also assume that the whole system is in an external, constant, homogeneous magnetic field B"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002355_j.jappmathmech.2012.03.010-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002355_j.jappmathmech.2012.03.010-Figure5-1.png",
        "caption": "Fig. 5.",
        "texts": [
            " It is noteworthy that the force is only present in the case of punches having corners with an unbounded contact pressure (in the case f a smooth punch = 0). This suggests that the solution to the paradox revealed is to be sought in the existence of several point tangential orces \u00b1, that are applied to the corners of the punch and compensate the tangential force formed within the contact area. These forces an be caused by the lateral action on the punch of the singularly deformed half-plane in the neighbourhoods of x = -a - 0 and x = b + 0 Fig. 5). . Determination of the corner tangential forces We will now determine the corner tangential forces \u00b1 by taking to the limit from a smooth punch to a punch with corners. We consider smooth punch L\u03b5, the shape of which is described by the function here g(x) is a smooth function. It is easy to observe that taking the limit \u2192 0 transforms the punch L\u03b5 into the punch L0 of shape y = (x) with corners (Fig. 6). We introduce the following notation We now define the tangential corner forces for the punch L0 using the formula (4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001538_978-3-642-34816-7_4-Figure4.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001538_978-3-642-34816-7_4-Figure4.4-1.png",
        "caption": "Fig. 4.4 Single inverted pendulum system",
        "texts": [
            " If the parameter values are chosen inappropriately, Gaussian function will not be effectively mapped, and RBF network will be invalid. According to the practical scope of x1 and x2; the parameters of ci and bi are designed as \u00bd 2 1 0 1 2 and 0.20; the initial weight value is chosen as zero. Adopting control law (4.6) and adaptive law (4.8), choosing Q \u00bc 500 0 0 500 , kd \u00bc 50; kp \u00bc 30; and \u03b3 \u00bc 1; 200: The results are shown in Figs. 4.2 and 4.3. The Simulink program of this example is chap4_1sim.mdl; the Matlab programs of the example are given in the Appendix. Consider a single inverted pendulum system as in Fig. 4.4. The dynamic equation is described as _x1 \u00bc x2 _x2 \u00bc f \u00f0x\u00de \u00fe g\u00f0x\u00deu where f \u00f0x\u00de \u00bc g sin x1 mlx2 2 cos x1 sin x1 \u00f0mc\u00fem\u00de= l 4 3= m cos2 x1 \u00f0mc\u00fem\u00de=\u00f0 \u00de ; g\u00f0x\u00de \u00bc cos x1 \u00f0mc\u00fem\u00de= l 4 3= m cos2x1 \u00f0mc\u00fem\u00de=\u00f0 \u00de ; x1 and x2 are angle and angle speed value, respectively; g \u00bc 9:8m=s2; mc \u00bc 1 kg is mass of cart; m \u00bc 0:1 kg is mass of the pendulum; l \u00bc 0:5m is the half length of the pendulum; and u is control input. Consider ideal position signal as yd\u00f0t\u00de \u00bc 0:1 sin t; the initial states are chosen as \u00bd\u03c0=60; 0 . RBF network structure is chosen as 2-5-1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.15-1.png",
        "caption": "FIGURE 1.15. Illustration of , , and rims and their geometry.",
        "texts": [
            " A ange has two bead seats providing radial support to the tire. The well is the middle part between the bead seats with su cient depth and width to enable the tire beads to be mounted and demounted on the rim. The rim hole or valve aperture is the hole or slot in the rim that accommodates the valve for tire in ation. There are two main rim shapes: 1 drop center rim ( ) and, 2 wide drop center rim ( ). The may also come with a hump. The humped may be called . Their cross sections are illustrated in Figure 1.15. Drop center ( ) rims usually are symmetric with a well between the bead seats. The well is built to make mounting and demounting the tire easier. The bead seats are around 5 deg tapered. Wide drop center rims ( ) are wider than rims and are built for low aspect ratio tires. The well of rims are shallower and wider. Today, most passenger cars are equipped with rims. The rims may be manufactured with a hump behind the bead seat area to prevent the bead from slipping down. A sample of rim numbering and its meaning is shown in Figure 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.22-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.22-1.png",
        "caption": "Figure 4.22 Parameters of bath lubrication",
        "texts": [
            " The factor of power losses in the mesh is determined by the following equation: HV \u00bc p 1 z1 \u00fe 1 z2 1 cos bb 1\u00fe e21 \u00fe e22 ea \u00f04:14\u00de For a bath lubrication, for a single gear pair, the hydraulic moment of power losses (in Nm) is determined as: TH \u00bc CSpC1e C2\u00f0v=vt0\u00de \u00f04:15\u00de where v is the reference circle peripheral speed of dived gear (in m/s), vt0 \u00bc 10 m=s is a constant, CSp is the factor of oil spraying and C1 and C2 are factors taking into account the influence of the gear width and the depth of their diving in the bath on power losses in idle motion (due to churning). The factor of oil spraying CSp\u00bc 1 for the arc w\u00bcp/2 which the involved oil has to pass until the mesh is shown in Figure 4.22a. For another possible value, w\u00bc 3p/2, Figure 4.22b, the factor of oil spraying is equal to: CSp \u00bc 4emax 3hC 1:5 2hC lh \u00f04:16\u00de where emax is the maximum of two values e1 and e2 of the diving depth of mated gears, hC is the height of the pitch point over the deeepest dive point and lh is the hydraulic length of the gear drive housing: lh \u00bc 4AK=O \u00f04:17\u00de where AK is the inside surface of the gear drive housing near the wheel gear and O is perimeter of that surface. The hydraulic length of the housing is commonly in the range lh\u00bc (500 . . . 1000 . . "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003648_icdma.2012.194-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003648_icdma.2012.194-Figure8-1.png",
        "caption": "Figure 8. 3d model",
        "texts": [],
        "surrounding_texts": [
            "It is the first thing to [10] appoint the CATIA tape, as shown in Fig. 2. There are two ways to visit the CATIA through the VB or VBA, as follows [10] 1) If the CATIA has been started, we just need to connect to the CATIA by the following program: : Dim CATIA as INFITF.Application Set CATIA = GetObject ( , \"CATIA. Application\") 2) If the CATIA hasn\u2019t been started, we need to start the CATIA at first by the following program: Dim CATIA as INFITF.Application Set CATIA = CreateObject (\"CAT IA. Application\") It\u2019s necessary to combine these two methods to ensure no mistakes when starting the CATIA. Namely: On Error Resume Next Dim CATIA as INFITF.Application Set CATIA = GetObject (, \"CATIA. Application\") If Err. Number <> 0 Then Set CATIA = CreateObject\"CATIA. Application\") CATIA.Visible = True B. The relationship of the basic parameters Transverse module: = Transverse pressure angle: = tan 1(tan cos ) Reference circle radius: = 2 Radiusofaddendum: = + Dedendum circle radius: = ( + ) Radiusofbasecircle: = cos Lead of helix: = 2 / tan Transition fillet radius: = (1 ) C. Establish the tooth profile In the course of modeling, the key problem is know how to get good tooth profile, the only way to ensure the worthy of the model and necessity of further study is to guarantee the exactness of the tooth profile. In the present study of involute gear modeling method, there are mainly two kinds of methods: 1) Tracing method Use the fog function and involute equations [11] The main idea of tracing method is the values of t. we can get some points belongs to the theoretical involute with different values of t, such as 0.1, 0.2, 0.25, 0.3, 0.35, 0.4... , each value corresponds one point. And then, we\u2019ll get the approximate involute through smooth connecting these discrete points with the spline curve. And the closer the values are taken, the close the obtained spline curve will to the theoretical involute, as shown in Fig. 3. after established the parameters: = = + 2) Direct method The so-called direct method is to directly generate involute with the involute equation. The specific process is: establish parallel curve 1 and parallel curve 2 respectively in plane yz and plane zx which corresponds to fog x and fog y with the parallel curve function of CATIA. And then, mix parallel curve 1 with parallel curve 2 for one curve through mixture function. Finally, project the mixed curve on plane xy to get the standard involute, as shown in Fig. 4. The involute will always have an endpoint located on (0, ) whether through tracing method or direct method, when the base circle fell on the lateral of dedendum circle, we need to make the endpoint of involute located inside of the base circle by extrapolation, avoid the endpoint located on (0, ). This is the only way to make sure that there are no mistakes when generate the transition fillet because of the involute do not intersect with the dedendum circle, in other words (the spur gear as example): > ( ) When = 20\u00b0 = 1 c = 0.25 . < 41.45 . That means when the number of teeth is less than or equal to 41, we need to do the extrapolation, and the different value of c will cause different situation. It\u2019s not necessary to do the extrapolation if the Radiusofbasecircle is less than the dedendum circle radius. This cause the ordinary method often results in an error, so we need to seek one way to avoid the extrapolation no matter what parameters we need. In this paper, the idea is to use the involute equation to gain the symmetrical involute relative to the original involute, namely: = + = We can get enough points on the involute through the circle program on taking the value of t in VB, and the finally spline curve will be very close to the theoretical involute as long as having enough points. It is 40 points that be taken on the involute to finish the spline curve, 20 points on original involute, and 20 points on symmetrical involute, as shown in Fig. 5. The final involute includes two involutes actually. And it is very convenient to use the same involute in the process of establishing the internal gear. In general, the method to create the involute not only can solve the problem of the extrapolation, but also can be a solution to the problem of internal and external involute gear. D. Modeling After getting the involute, we can directly do the clip and the symmetry on the involute, and finally generate a tooth profile, as shown in Fig. 6. We can get the gear slot through scanning resection the tooth profile along the spiral, and can all the gears through circle arraying the gear slot. E. Perfect the model It is necessary to perfect the 3d involute model generated by circle arraying, such as the spoke design for external gear, the keyslot design and the chamfering design, etc. All these details can be realized by VB command and finally can get the actual engineering 3 d model. IV. ENGINEERING EXAMPLE We can get the involute gear fast modeling software through the CATIA secondary development on involute gear with the VB program, the main window as shown in Fig. 7. Here, I will establish spur gear, inclined gear, internal gear and external gear model with the following parameter, = 50, = 2, = 20\u00b0, = 40, = 1, = 0.25 , = 70 , = 10 , = 30\u00b0 . The 3d models are shown in following figures. V. CONCLUSION This article has find one way which makes it possible for us to modeling both spur gear and inclined gear, both internal gear and external gear at the same time through the study on the method of involute gear parameterized modeling, and finally get the parameterized modeling software which allows us quickly and effectively complete the involute gear 3d modeling work as long as inputting the basic parameters. It can greatly improve the design efficiency and win the time for further design. VI. ACKNOWLEDGMENTS Here, we will thanks a lot to Hefei University of Technology, because the progress of the research is supported by the Doctors\u2019 Special Fund (2011HGBZ0928). We will also thanks to the national energy saving and new energy vehicles (863) project (2011AA11A236) and (2012AA112201 ). REFERENCES [1] Li Zisheng, Zhu Ying, Xiang Zhongfan. The technology of Secondary Development Based on CATIA[J]. Journal of Sichuan Industrial Institute, 2003. (in Chinese) [2] Long Feng, Fan Liuqun. The Research of the Secondary Development on CATIA V5[J]. Journal of Huaiyin Engineering, 2005. (in Chinese) [3] Cao Chunling, Fan Lili. The Involute Gear Parameterized Design Based on CATIA [J]. Lifting Transportation Machinery, 2008. (in Chinese) [4] Zhou Xian\u2019e, Lu Mowu, Zhao Haixing. The Study on CATIA Secondary Development Based on CAA [J]. Science and Technology Information, 2008. (in Chinese) [5] Su Hongjun, Wang Yongjin. The Study on CATIA V5 Secondary Development Based on CAA [J]. Machinery, 2008. (in Chinese) [6] Zheng Wenwei, Wu Kejian. Principle of Machinery[M]. Beijing: Higher Education Press, 1997. (in Chinese) [7] Wang Xiankui. Gear, Worm Gear and Worm, Spline Processing [M].Beijing: Mechanical Industry Press, 2009. (in Chinese) [8] Zhang Xunfu, Huang kang, Chen Ji. The Establishment of Involute Gear Dedendum Transition Curve Equation and the Accurate 3d Model[J]. Combination Machine Tools and Automatic Processing Technology, 2007. (in Chinese) [9] ChenJ i, Zhao Han, Huang kang. The Study on Parameterized Gear Precise CATIA Model Based on the Total Tooth Profile[J]. Combination Machine Tools and Automatic Processing Technology, 2009. (in Chinese) [10] Hu Ting, Wu Lijun. The Foundation of CATIA Secondary Development Technology[M]. Beijing: Electronic Industry Press, 2006. (in Chinese) [11] Xu Hongji, Tao Yanguang, Lei Guang. Gear Manual [S]. Beijing: Mechanical Industry Press,2000. (in Chinese) [12] Li Lizong. The Course of VB Program Design[M]. Tianjin: Nankai University Press, 2009. (in Chinese) [13] Ma Guoguang. Visual Basic Program Design [M]. Beijing: Tsinghua University Press, 2011. (in Chinese) [14] Wang Ping, Nie Weijiang. The Basic Course of Visual Basic 6.0 Program Design[M]. Beijing: Electronic Industry Press, 2012. (in Chinese)"
        ]
    },
    {
        "image_filename": "designv11_100_0003045_mias.2011.943100-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003045_mias.2011.943100-Figure8-1.png",
        "caption": "Figure 8 as",
        "texts": [
            " To confirm the adequacy of the emulated Epstein frame circuit model and provide a better illustration of the proposed evaluation procedure, a 300-W, three-phase, 60-Hz, 190.5-V synchronous switched-reluctance motor (SRM) with six stator poles and four rotor poles (6/4) is selected for assessment comparisons. As shown in Figure 7, the related 3-D structure and physical dimensions of this 6/4 synchronous SRM are illustrated. Both the stator and the rotor iron cores of this motor are assembled by the China Steel Corporation CSC 50CS470 laminated electromagnetic steel with its magnetic characteristics provided in Figure 8 and [7]. The related electrical and mechanical \u03b81 \u03b82 c d a b e a = 76 mm b = 15 mm \u03b81 = 26.5\u00b0 \u03b82 = 70\u00b0 c = 41.9 mm d = 37.4 mm e = 50 mm Air Gap = 0.5 mm 7 The 3-D structure and physical dimensions of a test 6/4 synchronous SRM. 30 IE E E IN D U S TR Y A P P LI C A TI O N S M A G A Z IN E J A N jF E B 2 0 1 2 W W W .I E E E .O R G /I A S parameters of the magnetic steel strips for performing the standard Epstein frame test are provided in Table 1. The Common Estimation Schemes Based on the iron loss information as provided by the manufacturer, if the motor operational flux density is at 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure1-1.png",
        "caption": "Figure 1. Genesis Front Multi-Link Suspension System.",
        "texts": [
            " So if only engineering plastics is used as the suspension parts, problems of degrading the performance and quality will happen. In the papers, the composite material suspension arm could be developed highlighting advantage of light weight of plastic and compensating for disadvantage of low strength of plastic. The concept of a composite material suspension arm is that with using a dual material of plastic and steel, plastic and steel is a technique for combining properties and large force is absorbed to steel and plastics compensates for steel. In the papers, such as a Fig 1 is development of a composite material suspension arm that are based on upper arm of genesis suspension system. A composite material suspension arm is developed with the same strength and stiffness compared with forged steel upper arm. And durability verification is proved the part and vehicle durability test based on Hyundai Motors specifications. The basic structure of composites suspension arm fixed to mold from steel press and plastic is injected into mold. However it is easily separated if the coefficient of friction, surface and material properties in the structure are different from other structure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003187_amm.483.382-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003187_amm.483.382-Figure5-1.png",
        "caption": "Figure 5(a) The main view of rotating brush Figure 5(b) The left view of rotating brush",
        "texts": [],
        "surrounding_texts": [
            "In recent years, as the scale of China's power grid construction has grown, the security of reusable equipment and supplies for power grid construction demands increasingly [1] . The wire rope used for tension wire of transmission line is a key instrument, its integrity is crucial for construction quality and safety of workers and equipment. Wire rope used for tension stringing work in relatively poor conditions, it easily produce wire rope wear, fatigue, broken wires, corrosion and deformation. According to information at home and abroad, comprehensive cleaning, safety testing and maintenance for a large number of wire rope has not been possible effective solution. To this end, according to the actual needs of production, Design and development of Wire rope automatic cleaning, detection and maintenance integration system has important practical significance, and it can produce good economic. The decontamination is the first procedure of clean module, the merits of equipment performance has an important impact."
        ]
    },
    {
        "image_filename": "designv11_100_0003474_amm.130-134.1205-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003474_amm.130-134.1205-Figure3-1.png",
        "caption": "Fig. 3 Mechanical structure of the system",
        "texts": [
            "34, University of Alberta, Edmonton, Canada-27/04/15,03:58:33) dkd rL VC V m SS \u22c5=\u22c5= \u03c0\u03b52 0 (1) The equation (1) can be obtained approximately when the capacitive sensor locates in centre of the measured hole. k is the equivalent coefficient for the instrument. Output voltage is linear with diameter of measured hole. The mm.01\u03c6 capacitive sensor is designed to satisfy smaller inner diameter measurement. And structure of the measuring head is shown in Fig.2: The measuring system composes of a main measuring bench, stepping motor control module, vision module and a two-channel capacitive micrometer (circuit processing module). The mechanical structure of the system is shown in Fig.3: In this automatic measuring system, accurate positioning is achieved by cooperating stepping motors with gratings respectively in X, Y, Z direction. Stepping motor is driven through PCI port by computer and the indication of grating is regarded as effective information to control the stepping motor. In this way, positioning accuracy of stepping motor is improved greatly. In order to eliminate error due to eccentricity of sensor and measured hole, firstly rough centering is conducted at end face of them by CCD vision and stepping motor to ensure there is no collision when putting measuring head into hole"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.10-1.png",
        "caption": "Figure 12.10 A moving body B in a global frame G .",
        "texts": [
            "214) I1I2I3\u03c8\u0308 sin2 \u03b8 = M\u03b8I3 (I1 \u2212 I2) cos \u03b8 cos \u03c8 sin \u03c8 + M\u03c8I1I2 sin \u03b8 \u2212 M\u03d5I2 (I1 \u2212 I3) sin \u03b8 cos \u03b8 sin2 \u03c8 \u2212 \u03d5\u03072 (2I1I2 (I1 \u2212 I2) + I2I3 (I2 \u2212 I3) +I3I1 (I3 \u2212 I1)) cos2 \u03b8 sin2 \u03b8 cos \u03c8 sin \u03c8 + \u03d5\u03072I1I2 (I1 \u2212 I2) cos \u03c8 sin \u03c8 \u2212 \u03b8\u03072 (I1I2 (I1 \u2212 I2)) sin2 \u03b8 cos \u03c8 sin \u03c8 + \u03d5\u0307\u03b8\u0307 [2I1I2 (I1 \u2212 I2) + I2I3 (I2 \u2212 I3) + I3I1 (I3 \u2212 I1)] cos2 \u03b8 sin \u03b8 cos2 \u03c8 + \u03d5\u0307\u03b8\u0307 [2I1I2 (I1 \u2212 I2) cos2 \u03c8 \u2212 I1I2(I1 \u2212 I2) \u2212 I1I2I3] sin \u03b8 + \u03d5\u0307\u03b8\u0307 [I1I2(I1 \u2212 I2) + I2I3(I2 \u2212 I3)] cos2 \u03b8 sin \u03b8 \u2212 \u03d5\u0307\u03c8\u0307[I2I3(I2 \u2212 I3) + I3I1(I3 \u2212 I1)] cos \u03b8 sin2 \u03b8 cos \u03c8 sin \u03c8 \u2212 \u03c8\u0307 \u03b8\u0307 [I2I3(I2 \u2212 I3) + I3I1(I3 \u2212 I1)] cos \u03b8 sin \u03b8 cos2 \u03c8 + \u03c8\u0307 \u03b8\u0307 [I2I3(I2 \u2212 I3) \u2212 I1I2I3] cos \u03b8 sin \u03b8 (12.215) If the rigid body is axisymmetric, I1 = I2 = I , then these equations will simplify to I \u03d5\u0308 = M\u03d5 + \u03d5\u0307\u03b8\u0307 (I3 \u2212 2I ) cot \u03b8 + \u03b8\u0307 \u03c8\u0307I3/sin \u03b8 (12.216) I \u03b8\u0308 = M\u03b8 + \u03d5\u03072(I \u2212 I3) cos \u03b8 sin \u03b8 \u2212 \u03d5\u0307\u03c8\u0307I3 sin \u03b8 (12.217) \u03c8\u0308I3 = M\u03c8 + \u2212\u03d5\u0308 (I3 \u2212 I ) cos \u03b8 + \u03b8\u0307 \u03c8\u0307I3 cot \u03b8 + 1 sin \u03b8 \u03d5\u0307\u03b8\u0307 ( I3 cos 2\u03b8 \u2212 2I cos2 \u03b8 ) (12.218) 12.3 RIGID-BODY TRANSLATIONAL DYNAMICS Figure 12.10 illustrates a moving body B in a global frame G. Assume that the body coordinate frame B is attached at the mass center C of the rigid body. The Newton equation of motion for the whole body in the global coordinate frame is GF = m GaB (12.219) This equation can be expressed in the body coordinate frame as BF = m B GaB + m B G\u03c9B \u00d7 BvB (12.220) \u23a1 \u23a3 Fx Fy Fz \u23a4 \u23a6 = \u23a1 \u23a2 \u23a3 m ( v\u0307x \u2212 (vy\u03c9z \u2212 vz\u03c9y )) m ( v\u0307y \u2212 (vz\u03c9x \u2212 vx\u03c9z) ) m ( v\u0307z \u2212 (vx\u03c9y \u2212 vy\u03c9x )) \u23a4 \u23a5 \u23a6 (12.221) In these equations, GaB is the global acceleration vector of the body at C, m is the total mass of the body, and F is the resultant of the external forces acted on the body at C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003846_s1068798x11030129-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003846_s1068798x11030129-Figure2-1.png",
        "caption": "Fig. 2. Engagement of proposed gear pair in initial position (a) and after rotation of gear 1 by an angle \u03b31 (b).",
        "texts": [
            ") Synthesis of the transition curves at the tooth base is divided into two stages. First, we calculate the tran sition curve of the tooth base such that gear rotation is associated with successive contact of the base with the tooth tip of the conjugate gear (as in gear knurling, for example). To this end, we construct curves (roulettes) associated with initial gear circumferences that roll over one another. The roulettes are described by the normals passing though the instantaneous poles of rotation [3]. In Fig. 2, we show such a gapless engage ment (without a radial gap) of two involute gears with asymmetric tooth profiles in contact with the working profiles. In Fig. 2a, we show the initial gear position; gear 1 is assumed to be the generatrix of the transition curve of the base of gear 2. We employ inversion of the motion. Then gear 2 is immobile, while gear 1 (radius of initial circumference will roll over the initial circumference (radius of gear 2. If the arcs P and P are equal, gear 1 turns around its axis O1 by an angle \u03b31 and around axis O2 by an angle \u03b32. The new position of gear 1 is shown in Fig. 2b. The contact nor mal AB passes through the new rotation pole P '. \u03b1W* rW1 ) rW2 ) P1' P2' The mobile coordinate system x 'O1y ' is associated with gear 1; its origin O1 is at the center of gear 1. The y ' axis is aligned with the axis of the channel in the working tooth profile of gear 1. Point A at the tooth tip of gear 1 has the coordinates x ' and y ' in coordinate system x 'O1y '. The trajectory of this point in immobile coordinate system xO2y will represent the desired pro DOI: 10.3103/S1068798X11030129 RUSSIAN ENGINEERING RESEARCH Vol. 31 No. 3 2011 TRANSITION CURVE OF INVOLUTE GEAR TEETH WITH ASYMMETRIC PROFILES 201 file of the tooth channel in gear 2 for the first stage of synthesis. In geometric calculation of the asymmetric engagement profile, the coordinates of point A are determined as xy and yy in coordinate system xyO1yy (Fig. 2a) [2]. Transition to coordinate system x 'O1y ' may be based on the familiar formulas (1) For the given transmission, \u03b8 = 4.5\u00b0, \u03d5 = 12.8396\u00b0, xy = 158.5 mm, and yy = \u201336.0 mm [2]. From Eq. (1), we obtain the coordinates of the tooth tip (point A): x ' = 12.874 mm; y ' = \u2013161.702 mm. These coordinates remain unchanged with rotation of gear 1, of course. Since arcs P and P are equal, then \u03b31 = \u03b32(z1/z2). In the present case, when z1 = z2, we find that \u03b31 = \u03b32. Using the dimensional chain corresponding to Fig. 2b, we obtain formulas for the coordinates of points of the transition curve in the first stage of syn thesis, when it is formed by the tooth tip of the conju gate gear Here the argument is the rotary angle \u03b32. Given that z1 = z2, we obtain (2) x' xy \u03b8 \u03d5+( )sin yy \u03b8 \u03d5+( );cos+= y ' xy \u03b8 \u03d5+( )cos\u2013 yy \u03b8 \u03d5+( ).sin+= \u23ad \u23ac \u23ab P1' P2' xA aW \u03b32sin y ' \u03b32 z1 z2+( ) z1 sin x ' \u03b32 z1 z2+( ) z1 cos ;+ += yA aW \u03b32cos y ' \u03b32 z2 z1+( ) z1 cos x' \u03b32 z1 z2+( ) z1 sin .\u2013+= \u23ad \u23aa \u23aa \u23ac \u23aa \u23aa \u23ab xA aW \u03b32sin y ' 2\u03b32sin x ' 2\u03b32;cos+ += yA aW \u03b32cos y ' 2cos \u03b32 x ' 2sin \u03b32"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003133_0954406211417228-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003133_0954406211417228-Figure2-1.png",
        "caption": "Fig. 2 Location of pads on piston",
        "texts": [
            " In the manner of Cameron [6], for wide bearings the constants n and k would be n\u00bc 3 and k \u00bc 12 wB3 \u00f04\u00de where B is the length of the bearing, the dynamic viscosity of the oil, and w the width of the bearing. The value of k, using the dimensions of the bearings under consideration, is in the region of (2! 8) 10 9, and this range of values was used in the mathematical model of the engine. The piston skirt has been divided into 32 identical pads, each occupying an angle of 45 either side of the centre, as shown in Fig. 2. This number of pads was chosen because no noticeable difference was detected in the results when the number of pads was increased from 16 to 32. The force acting on each of the pads could arise from both elastic and hydrodynamic forces and is determined as follows. Referring to Fig. 11(a) in Appendix 2, the gap between the piston skirt and the cylinder wall at pad YPn is given by EPgap \u00bc GAP xp \u00fe XCYL \u00fe GX \u00fe \u00f0Pa YPn\u00de \u00f05\u00de If EPgap< 0, then EPn \u00bc EPgapKp \u00f06\u00de and HPn \u00bc 0 \u00f07\u00de where EPn is the elastic contact force, HPn the squeeze film force, and Kp the contact stiffness between the piston and cylinder wall"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.26-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.26-1.png",
        "caption": "Figure 10.26 Four commonly used IPM rotor configurations: (a) circumferential-type magnets suitable for brushless DC or synchronous motor; (b) circumferential-type magnets for line-start synchronous motor; (c) rectangular slots IPM motor; and (d) V-type slots IPM motor",
        "texts": [
            " It is worth pointing out that although there is a reluctance torque component, it does not necessarily mean an IPM motor will have a higher torque rating than a SPM motor for the same size and same amount of magnetic material used. This is because, in IPM motors, in order to keep the integrity of the rotor laminations, there are so-called \u201cmagnetic bridges\u201d that will have leakage magnetic flux. So for the same amount of magnet material used, a SPM motor will always have higher total flux. There are many different configurations for IPM motors as shown in Figure 10.26. The no-load magnetic field of PM machines is shown in Figure 10.27. When the rotor is driven by an external source (such as an engine), the rotating magnetic field will generate three-phase voltage in the three-phase windings. This is the generator mode operation of the PM machine. When operated as a motor, the three-phase windings, similar to those of an induction motor, are supplied with either a trapezoidal form of current (brushless DC) or sinusoidal current (synchronous AC). These currents generate a magnetic field that is rotating at the same speed as the rotor, or synchronous speed",
            " However, temperature effects, as shown in Figure 10.33, must be taken into consideration when designing a PM motor. At room temperature (25 \u25e6C), the demagnetizing curve of a cuboid REPM can be represented by m = r \u2212 Fm \u00b7 r/Fc = r \u2212 Fm/RM (10.76) where r and Fc are the residual flux and magnetomotive force (mmf) of each pole respectively, and RM is the reluctance of the magnet, which is the reciprocal of magnet permeance \u03bbM : RM = 1/\u03bbM = Fc/ r (10.77) For parallel or circumferentially magnetized poles as shown in Figure 10.26a,b, r = 2BrAm, Fc = lmHc (10.78) while for series or radially magnetized poles as shown in Figure 10.26c,d, r = BrAm, Fc = 2lmHc (10.79) where Br and Hc are the remanence and coercive force of the magnets respectively, lm is the length of the magnet, and Am is the cross-sectional area. Thus Am = bmlf e (10.80) where lf e is the length of magnet along the shaft direction and usually equal to the rotor lamination stack length. At the operating temperature, the above parameters are replaced by their respective values. It is possible that the demagnetizing curve becomes nonlinear at the operating temperature"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002894_amr.837.88-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002894_amr.837.88-Figure2-1.png",
        "caption": "Fig. 2",
        "texts": [
            " In the case when in the point B acts on the distance OB an exterior force 0F , too, then in relation ( )[ ] ( ) 11 1112 sin2 sin2cos2 \u03b8 \u03b8\u03b8 \u03b1 \u22c5\u22c5+ \u22c5\u22c5+\u2212\u22c5+ = lQF lQFQM tg we replace F by 220 xmF \u2212 . In the situation that it results 0N < from relation ( )[ ] ( ) 11 1112 sin2 sin2cos2 \u03b8 \u03b8\u03b8 \u03b1 \u22c5\u22c5+ \u22c5\u22c5+\u2212\u22c5+ = lQF lQFQM tg one deduces the angle. ( )[ ] ( ) 11 1112 sinlQF2 lsinQF2cosQM2 arctg \u03b8 \u03b8\u03b8 \u03c0\u03b1 \u22c5\u22c5+ \u22c5+\u2212+ += . (5) That means when reaction force N passing through zero position a jumping is produced. We consider that the crank 1OO of the mechanism drawn in Figure. 2 rotates with constant angular speed \u03c9 and that in the point B acts in the direction OB a force 0F (to keep the clearance). Denoting by ii yx , the co-ordinates of the weight centers for the elements ( ) ( ) ( )3,2,1 , by i\u03b8 the positional angles, by im the masses and by iJ the central moments, and by 4J the inertial moment with respect to the point O of the crank, one obtains the expression of the kinetic energy. ( ) ( )[ ]2 4 2 33 2 3 2 33 2 22 2 11 2 1 2 11 2 1 \u03c9\u03b8\u03b8 JJyxmxmJyxmEC +++++++= . (6) where 0;0 33 == Jm and the generalized forces 02 FQ = "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.6-1.png",
        "caption": "Figure 2.6. The circuits of the synchronous generator: (a) rotor circuit; (b) stator circuit.",
        "texts": [
            " (ii) The stator slots cause no appreciable variation of the rotor inductances with rotor position. (iii) Magnetic hysteresis is negligible. (iv) Magnetic saturation effects are negligible; the machine equations will be developed first by assuming linear flux\u2013current relationships. In order to study the power system operating conditions, the synchronous generators are represented as a number of equivalent windings, magnetic coupled and rotating. From Figure 2.1, the following circuits are identified (Figure 2.6): Stator Circuits. The three stator windings a\u2013a0, b\u2013b0, and c\u2013c0, distributed 120 apart in trigonometric rotation direction. The voltages at the stator winding terminals s Ce Cm Ms D \u03b4\u03c9 \u03c9 + 1+ are va, vb, and vc, and the currents are ia, ib, and ic. Using the generator convention, the stator current is considered positive when it is out of the machine. Rotor Circuits. The field (excitation) winding f\u2013f0 is taken as the direct axis, which is simply called d-axis, and the quadrature axis, which is 90 ahead of the direct axis in the rotational direction, is called q-axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003046_cdc.2011.6160703-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003046_cdc.2011.6160703-Figure1-1.png",
        "caption": "Fig. 1. Inertia Wheel Pendulum (IWP)",
        "texts": [
            " We skip the controller design and we assume that a control law u(x) is given for each application such that the closed loop dynamics x\u0307 = [ \u03b7\u0307 \u00b5\u0307 ]T = f (\u03b7 ,\u00b5 ,u(\u00b5 ,\u03b7)) (25) is (locally) asymptotically stable. Due to the technical restrictions the system variables \u03b7 are not available as measurement and one has to replace \u03b7 by the estimated values \u03b7\u0302 in the control law. Some simulation results are included for an intensive study of the modified control law u(\u00b5 , \u03b7\u0302) and estimation error e = \u03b7 \u2212 \u03b7\u0302 . Here the models, the parameters and the control laws are taken from the literature and we focus only on the design procedure for the observer. The inertia wheel pendulum, see Fig. 1 is also known as reaction wheel pendulum [13]. It is a 2-dof mechanical example, which has been investigated with many different control design methods and the IWP has become a benchmark for the control of underactuated systems. The IWP is just a version of the well-known inverted pendulum, which uses a gyroscopic actuation. We consider a mathematical model Mq\u0308e +G(qe)+D(q\u0307e) = Me qT e = [ \u03d51 \u03d52 ] , (26) with a constant inertia matrix M \u2208 R 2\u00d72, G(qe) resulting from the gravity, D(q\u0307e) for the dissipative terms and the actuator torque Me"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001685_978-1-4302-4387-8_6-Figure4-14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001685_978-1-4302-4387-8_6-Figure4-14-1.png",
        "caption": "Figure 4-14. Kyle the Robot is a work-in-progress who requires a wide range of specialized tools and",
        "texts": [
            " The basic framework is made of hand -tooled aluminum. The large wheels use O-rings for tires and are driven by 24V gear motors. The 24V supply is made up of two 12V, 5Ah rechargeable lead-acid batteries. Ah stands for amp hours, and represents the capacity of the battery. In this case, it can deliver an amp of current for five hours, or five amps for one hour. The custom PCB holds the electronics for the drive motors and will be controlled by a separate microcontroller that has yet to be added. See Figure 4-14. 103 CHAPTER 4 . A PORTABLE MINI-LAB 104 components. Some new components are beingfitted, including dedicated battery voltmeters,fuses, a circuit breaker, and a master power switch. Once the sharp corners are filed off, Kyle will be able to roll around the house, lookingfor chores to do. As we come to the end of this chapter, you should have some good ideas bouncing around in your head about how you want to pack your mobile electronics lab, what tools and components you absolutely must have with you, and how you want them organized"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002199_978-1-4419-9323-6_4-Figure4.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002199_978-1-4419-9323-6_4-Figure4.2-1.png",
        "caption": "Fig. 4.2 Development of a per-phase equivalent circuit of a DFIG",
        "texts": [
            " Note that isd and isq in the GSC context refer to the d-axis and q-axis currents respectively in the grid side PWM converter, and not the stator currents. The per-phase equivalent circuit of a DFIG corresponding to a given slip, s, can be derived from the basic principles of operation of induction machines and transformers. Such an equivalent circuit is quite useful for steady-state analysis with phasors. In order to be consistent with the dynamic models introduced in subsequent sections as well as with models used in the literature, the current directions in the equivalent circuits are defined based on the motor convention. Figure 4.2 illustrates the development of a single-phase equivalent circuit, with the rotor impedances referred to the rotor side and the stator quantities referred to the stator side, coupled by the equivalent model of the rotating transformer. Rs and Rr are the physical resistances of the stator and rotor windings, respectively, Lls and Llr are the leakage inductances of the stator and rotor windings, and Lm is the magnetizing inductance referred to the stator side. Vs is the grid voltage applied at the stator and Vr is the rotor voltage applied by RSC, both represented as phasors; Is and Ir are the stator and rotor current phasors, respectively, as indicated",
            " The stator is stationary, and therefore, the relative velocity with any field quantity (which rotates atxs) is xs, while the rotor rotates at a velocity xr and its relative velocity with respect to any field quantity is xs xr \u00bc sxs. Therefore, for example, the magnitudes of stator side voltages when referred to the rotor side are multiplied by s. In addition, if the physical turns ratio between the stator and rotor windings is Nr/Ns = n, then the rotor voltage is scaled by an additional factor of n. So, the primary voltages when referred to the rotor side are multiplied by the factor ns as indicated in Fig. 4.2b. Unlike the voltage magnitudes, the current magnitudes are not scaled by s when moving across the rotating transformer. This is because the currents drawn from the stator of an induction machine to cancel the MMF produced by the rotor currents do not depend on the rotation or relative velocity of the individual windings. Only the ampere-turns in the stator windings are required to be equal to the ampere-turns in the rotor windings, with the assumption of the leakage inductance being negligible compared to the magnetizing inductance. Therefore, for example, when referring any current from the stator side to the rotor side, the current magnitude is multiplied by the physical turns ratio n alone. This is illustrated in Fig. 4.2b with the controlled current source at the stator side, which is equal to n Ir. It is important to remember that the frequency of the current is scaled by s. Since only the voltage is scaled by s and not the current, the impedances are scaled by s when transferring from stator to rotor (and not by s2 as is the case for physical turns ratio), and 1/s when transferring from rotor to stator. In a transformer with stationary primary and secondary windings the power processed by the primary and secondary windings are equal, which is consistent with the scaling of the voltage by the turns ratio n and current by 1/n when moving from the primary to the secondary. However, in an induction machine the power processed by the stator and the rotor windings are not equal as can be inferred from the scaling of only the voltage by s and not the current, and the value of power (active and reactive) processed by the two controlled sources not being equal. The power processed by the stator winding at the air gap Pg (which does not include stator resistive loss) and the power processed by the rotor winding at the air gap Pr are related by (4.1) as seen from Fig. 4.2b. Pr \u00bc s Pg \u00f04:1\u00de The difference in the active power processed between the two windings is the mechanical power, absorbed or sourced, depending on operation as a motor or generator, respectively, and is given in (4.2) and illustrated in Fig. 4.2b (motor convention). This is the power required to maintain the rotor at a velocity of xr at a torque of Tem (again positive or negative depending on operating mode). Pg Pr \u00bc 1 s\u00f0 \u00dePg \u00bc 1 s s Pr \u00bc Pmech \u00bc xrTem \u00f04:2\u00de A similar expression can be derived for the reactive power relationship also, as shown in (4.3), with the rotating transformer structure scaling the reactive power by s. This is an important relationship as it shows that the reactive power at the stator can be controlled by a significantly smaller reactive power injection at the rotor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure5-1.png",
        "caption": "Fig. 5 Details of the guiding track",
        "texts": [
            " An invention represents this kind of quick connection is about a medical device connector [8]. It relates to a connection site for a medical device having a neck element with at least one guiding track. The guiding track has a lock edge for cooperative engagement with a lock protrusion of a second medical device. Fig.4 shows a medical device 1 in the form of a medical device connector 1 for connecting two medical devices. The connection site 4 comprises a neck element 5 having three guiding tracks 6 for receiving lock protrusions 3 piercing device 2. As is shown in Fig.5, the guiding track 6 exhibits Lform, comprising a first vertical section, the guiding track 6 comprises the locking edge, or barrier section, which the lock protrusions 3 of the piercing device 2 are intended to cooperate with during assembly. During the insertion, the lock protrusions 3 of the piercing device 2 slide in the vertical section of the guiding track 6. The structure represented above can make a turning motion to connect or disconnect the male and female sides. While designing this kind of mechanical quick connection for medical devices, the angle C (in Fig.5) is the key point to make sure the structure works. A. Stricture Scheme A new kind of quick connector was designed based on the structure of a push-push type latch mechanism and electronic device thereof [9]. This machine was designed for a connection of a function bed and a devise which hold a tubular shape medical device, a catheter [10] for example. Fig.6 shows a full view of the connector. It includes a holding device which is the male part of the connection, and the locking part links with the function bed, which is the female part of the connection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002691_imece2012-87321-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002691_imece2012-87321-Figure7-1.png",
        "caption": "Figure 7. NREL GRC TURBINE DRIVE TRAIN SET-UP",
        "texts": [
            " The main assumption in the market is that this system only constrains the gearbox in those DOFs which should be constrained, while allowing the gearbox to move freely in all other DOFs, which should avoid unfavorable loading conditions. Model description Investigated system A full nacelle model including main shaft, gearbox and generator is needed to investigate the influence of the drive train layout and mounting conditions. It is opted to start from the wind turbine available in the Gearbox Reliability Collaborative (GRC) of NREL [2, 3], since there has been extensive validation on the system. A model of the GRC drive train is shown in Figure 7. The drive train features a TPM configuration typical of modular MW class wind turbine drive trains today. A single spherical roller main bearing and two elastomeric trunnions support the gearbox and main shaft assembly. The 2 speed synchronous generator mounts to the end of the bedplate and flexibly couples to the gearbox. The gearbox has a three planet low speed stage followed by two parallel stages with an overall ratio of 1:81.491. A flexible multibody model of the GRC drive train was built and is shown in Figure 7. The model used in this 4 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76595/ on 03/22/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Figure 8. THREE POINT MOUNTING CONFIGURATION MODEL Figure 9. FOUR POINT MOUNTING CONFIGURATION MODEL paper was defined according to the validated modeling described in [5] and bearing and gear forces during constant gearbox loading were validated against other models used in the GRC to improve confidence in the results"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.42-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.42-1.png",
        "caption": "Figure 2.42. Loading capability chart of a synchronous generator [30].",
        "texts": [
            ", hydrogen) or air temperature (which is variable with the season, e.g., summer or winter). In order to develop the capability curves of the synchronous generator we start from the phasor diagram of the synchronous machine under balanced steady state from Figure 2.16, in which we neglect the resistance since it is much smaller then the reactance and considering the round rotor generator, with XS \u00bc Xd \u00bc Xq. The saturation is also neglected as it involves negligible differences. In terms of powers, we obtain the so-called loading capability chart (Figure 2.42) [30]. In the following we define the three design limits of the generator that restrict provision of reactive power. (i) Armature Current Limit. The current produced by the generator results in heating of the stator windings and significant damages if limiting actions are not taken. For this reason, the value of the current carried in the stator winding must be limited to a certain value. In the P\u2013Q plane, this limit is a circle with center in the origin and radius equal to the rated apparent power (Figure 2.42). The complex output power is given by the expression _S \u00bc P\u00fe jQ \u00bc _V g _I g \u00bc VgIg cos wg \u00fe j sin wg E Z I V (a) (b) I Y E= Y\u2033 I V\u2033\u2033\u2033 \u2033 \u2033 Assuming nominal terminal voltage and armature current as well as nominal cooling conditions, the output apparent power is a combination of active and reactive powers. At the limit, more reactive power may require reduction of the active power and vice versa. (ii) Field Current Limit. The excitation current carried in the rotor winding causes also heating and energy losses and therefore a second limit is defined. In the plane P\u2013Q, this limit is given by the arc of the circle with origin in ( V2 g=XS) and radius (VgEq=XS). The origin of this circle and, therefore, the length of the radius depends on the size of the generator reactance. This circle is obtained by multiplying the phasors of Vg, Eq\u00f0\u00bc XmdIf \u00de, and XSIg from Figure 2.16 with Vg=XS. Therefore, the output powers can be expressed as P \u00bc VgIg cos wg \u00bc EqVg XS sin di (2.198) Q \u00bc VgIg sin wg \u00bc EqUg XS cos di V2 g XS (2.199) As seen in Figure 2.42, the field current limit is more restrictive then the armature current limit and the two curves intersect at the point N. In some cases, the point N is located on the prime mover limit and the field current limit is the only restriction for the generated reactive power. (iii) End Region Heating Limit. When operated in underexcitation the end-turn leakage flux, as illustrated in Figure 2.43, enters axially (perpendicular) to the RotorStator winding end turns Stator Retaining solid/iron ring Eddy currents Figure 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003670_amr.479-481.670-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003670_amr.479-481.670-Figure1-1.png",
        "caption": "Figure 1 shows the structure of typical wheel driving system with exterior rotor motor. The rotor in motor and the shell are connected to the wheel rim through bolt, while the stator is connected to the suspension pin and the fixed adapter. The brake disc is fixed to motor shell and the brake caliper connected to fixed adapter, finally the suspension pin is fixed to suspension assembly [3].",
        "texts": [],
        "surrounding_texts": [
            "Electric vehicle has become the mainstream of development of new energy vehicle in worldwide. The wheel hub motor based electric vehicle has compact structure, high utilization rate in interior space and great steering stability, but the integration design of wheel driving system is very complex, and thus the load and mechanical characteristics of main parts in system will be changed, especially the wheel hub bearing[1,2]. The preliminary analysis shows that, there are several structural parameters in wheel driving system, which will influence the mechanical characteristics of wheel hub bearing, thus ultimately affect system\u2019s structural design. Therefore, based on the development of a certain wheel driving system, this paper has analyzed the structural parameters and the influence of load on wheel hub bearing, in order to reduce the force on bearing as much as possible. Structure of Wheel Driving System"
        ]
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.104-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.104-1.png",
        "caption": "Fig. 2.104 4RRRRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology R||R||R\\R||Pa",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure9-1.png",
        "caption": "Figure 9. Penetration plot at RSFB region",
        "texts": [
            " Figure 6 and Figure 7 shows the comparison of displacements and stresses respectively without and with sub-modeling. Displacement and stress results with sub-modeling is compared with full model linear static analysis and observed the same results. This has given confidence that submodeling is properly done. FE model should be modeled without any penetration. This is a pre-requisite to carry out contact analysis otherwise it will lead to unrealistic results. Figure 8 shows the super element at RSFB region and Figure 9 shows the Penetration plot at RSFB region. It is observed that bottom region of the suspension mounting bracket is penetrating inside the external flitch. Figure 10 shows the comparison of stresses with and without penetration at RSFB region. It is observed from Figure 10 that unrealistic stress occurs at the location where penetration happens. Table 1 shows the comparison of stresses and displacements with and without penetration. It is observed from Table 1 that there is a significant difference in stresses"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002563_ijmms.2013.052783-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002563_ijmms.2013.052783-Figure8-1.png",
        "caption": "Figure 8 Following a leader robot (continued)",
        "texts": [
            " First, we define the primary triangle to determine the common coordinate. Then, the robots that compose the primary triangle are given the angle error or gap (\u03c6) from the apex of IET as shown in Figure 7. Then the robots rotate around the origin of the common coordinate, because they move to new destinations (P1, P2, P3) so that the rotational error may become zero. For example, consider the case where there is a remotely controlled leader robot in a formation. When the leader robot (i.e., Robot 1) moves as shown in Figure 8, the desired positions for other robots are altered so that they can form a desired formation around the leader robot. In other words, if Robot 1 moves arbitrary direction, then the other robots can move to keep the equilateral triangle formation. In this section, several simulation experiments are planned to check that any formation construction and its motion are implemented by using the IET algorithm. When robots move to closely approach each other, a logical rule of collision avoidance is assumed so that the robot turns to the left 90 [deg]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003420_s0218625x13500431-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003420_s0218625x13500431-Figure1-1.png",
        "caption": "Fig. 1. Geometrical model of asymmetry groove texture.",
        "texts": [
            "15,16 In order to further understand the lubrication mechanism of asymmetric texture and optimize the surface texturing model, a two-dimensional computational \u00b0uid dynamics (CFD) model based on Navier Stokes equations was adopted to simulate the hydrodynamic load-bearing capacity of asymmetric surface texture under the state of \u00b0uid lubrication in the present paper. Besides, it also conducted a detailed analysis of the in\u00b0uence of the changes in asymmetric parameter H and the Reynolds number Re on the load-bearing capacity of oil \u00aflm. In this work the in\u00b0uence of the asymmetry of single groove texture in the in\u00afnite surface texture on the load-bearing capacity of oil \u00aflm under reciprocating conditions was investigated. Figure 1(a) showed the three-dimensional schematic view of an in\u00afnite textured surface. The simpli\u00afcation of two-dimensional research model of single groove texture unit was obtained [Fig. 1(b)] because the distance between the adjacent asymmetry groove textures was much larger than the width of groove. The ratio H \u00bc h1=h2 in the groove texture [Fig. 1(b)] was used to measure its non-symmetry. In Fig. 1(b), the upper wall was moving while the lower wall was stationary. The unit length lx was 1e-03m, and the width d was 4e-04m. In addition, the viscosity and density of the lubricating oil were respectively set at 0.01Pa s and 1000 kg/m3. The lubricate \u00aflm thickness h\u00f0x\u00de was the gap between the upper and bottom planes. The minimum value of \u00aflm thickness was 3e-05m, and the thickness equation was shown as follows: h\u00f0x\u00de \u00bc Max\u00bdh\u00f0x\u00de \u00bc hout \u00fe h2 hinlet \u00fe h1 ( lx d 2 < x < lx \u00fe d 2 Min\u00bdh\u00f0x\u00de \u00bc hout x > lx \u00fe d 2 ;H \u00bc h1 h2 1 Min\u00bdh\u00f0x\u00de \u00bc hinlet x < lx d 2 ;H \u00bc h1 h2 1 8>>>>< >>>>: \u00f01\u00de 3",
            " To facilitate the result displaying and data analysis, the following dimensionless parameters were de\u00afned: X \u00bc x=lx; Y \u00bc y=h U \u00bc u=u0; V \u00bc v=v0 \u00f05\u00de \u00bc = 0; \u00bc = 0; P \u00bc p=p0: In the above equation, x and y were coordinate vectors, u0 and v0 were the characteristic velocity along the x- and y-direction, 0 and 0 were the characteristic density and dynamic viscosity of the lubricant, p0 was the characteristic pressure and lx was the characteristic length along the x-direction. The dimensionless Navier Stokes equation and the continuity equation were shown as follows: x-direction U @U @X \u00fe V @V @Y \u00bc 1 Re @P @X 1 \u00fe 1 Re @ 2U @X 2 \u00fe @ 2U @Y 2 ; \u00f06\u00de y-direction U @V @X \u00fe V @V @Y \u00bc 1 Re @P @Y 1 \u00fe 1 Re @ 2V @X 2 \u00fe @ 2V @Y 2 : \u00f07\u00de The dimensionless continuity equation: @U @X \u00fe @V @Y \u00bc 0: \u00f08\u00de 3.2. Computational domain and boundary conditions In order to simulate the pressure \u00afeld with reasonable boundary conditions in Fig. 1, the computational domain was established in Fig. 2. The district 1 in Fig. 2 showed the research model as shown in Fig. 1(b). District 1 and 2 were symmetric with respect to the straight line e and regarded as an integral unit, which can simply make use of the CFD model to simulate the relevant data of the entire \u00b0ow \u00afeld. What was studied in this paper was the in\u00b0uence of the asymmetry of the single groove texture in the in\u00afnite surface texture on the load-bearing capacity, as shown in Fig. 1. Therefore, the impact on the \u00b0ow \u00afeld in district 2 caused by district 1 could be ignored. The relevant data was obtained from the geometric model in Fig. 2, and then the data of district 1 was extracted as the approximate data of the internal \u00b0ow \u00afeld of single groove texture, as shown in Fig. 1(b), aiming to distinguish the in\u00b0uence of asymmetry of the single groove texture on the load-bearing capacity. Boundary conditions were speci\u00afed for the computational domain shown in Fig. 2. The upper wall moved along the x-direction at a velocity of U and noslip. The lower wall was stationary, and the inlet and outlet were set to the periodic boundary. The traditional \u00afnite volume method was used. The second order upwind scheme was used for the discretization of the momentum. The SIMPLEC method was used to couple the pressure and velocity",
            " Therefore, the average value of the pressure distribution of the upper wall with di\u00aeerent H was closer to that on the non-textured surface. The above analysis showed that the in\u00b0uence of H on the load-bearing capacity gradually decrease with the increase of Re. Based on the above analysis, the suitable asymmetric parameter H could promote the load-bearing capacity and improve the hydrodynamic lubrication at a low Reynolds number Re. But this in\u00b0uence become weakened with the increase of Re. This was mainly because of the smallerHmeans the greater inlet \u00aflm thickness relative to that of outlet [Fig. 1(b)], that was to say, a convergent gap was formed between the upper and bottom surfaces, and thereby generating an additional normal pressure to improve load-bearing capacity.17 When Re was increasing, the inertia term of \u00b0uid \u00b0ow gradually played a major role, and the pressure of \u00b0uid \u00b0ow caused by this was much greater than the dynamic pressure caused in the convergent gap. (c) (d) Fig. 3. (Continued) 1350043-5 Su rf . R ev . L et t. 20 13 .2 0. D ow nl oa de d fr om w w w .w or ld sc ie nt if ic "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002563_ijmms.2013.052783-Figure18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002563_ijmms.2013.052783-Figure18-1.png",
        "caption": "Figure 18 Symbol of simulation parameters (see online version for colours)",
        "texts": [
            " 3 Receiving the wireless signal and ultrasonic sound: Assuming that the wireless signal can approximately arrive at Robot 1 with 0 [s], the distance between the sensor and Robot 2 can be calculated with a counted time, which is the time until the ultrasonic sound arrives after receiving the wireless signal (see Figure 16). 4 Estimation of relative position: The relative position of Robot 2 (x2, y2) is estimated by figuring out the coordinate point where three circles, whose radius is the distance to Robot 2, cross each other (see Figure 17). In this section, we evaluate the accuracy of estimation of a relative position of a robot through two simulation experiments. Figure 18 shows the experimental environment. Robot 1 is a measuring robot, and the origin of the coordinate is defined as the centre of Robot 1, where the x-directional coordinate is just to be a line connected with the origin and Sensor 1. Although Robot 1 estimates the position of Robot 2 (x2, y2) based on the distance between sensors and Robot 2, the measured value includes noise up to \u00b1 1 [%]. The accuracy of estimation is evaluated for cases where the distance between sensors, L, or the distance between robots, D, changes independently"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002395_epepemc.2012.6397388-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002395_epepemc.2012.6397388-Figure10-1.png",
        "caption": "Fig. 10. Different configurations of rotor faults tested (BRBs marked black).",
        "texts": [
            " Laboratory Setup The proposed method for detecting the rotor fault has been tested on an experimental set-up, using a machine whose data are listed in the appendix. For this purpose, identical rotors (in theory) have been separately commissioned in order to test several diagnostic approaches. At the beginning, initial measurements were performed on all rotors in a healthy state in order to take account for all manufacturing imperfections that can potentially influence future diagnostics. After that, different number and combination of holes were drilled in rotor bars, at a junction with a ring, as shown in Fig. 10. B. Experimental results Measurements were performed at several operating points combining different speeds (750 rpm \u2013 Fig. 11 and 1500 rpm \u2013 Fig. 12) and load (0 Nm, 15 Nm, and 30 Nm). Please note that the phase shift is owed to visualization of different curves measured separately and gathered on a same oscillogram. Nevertheless, since all measurements were performed at the same load, oscillation frequency for all faults is equal, as explained in the introduction. An obvious conclusion that can be made from the results is that the amplitude of the oscillations strongly depends on the degree and type of the fault"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002764_0976-8580.99296-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002764_0976-8580.99296-Figure4-1.png",
        "caption": "Figure 4: Deformable flat contact model",
        "texts": [
            " For a higher strain it will vary with the strain, and depending on the material it could be any value from zero, up to bigger than the elastic modulus. A wide range of values of the tangent modulus are taken to get a fair idea about the effect of this in different materials, the hardening parameter, and the area of contact. The tangent modulus is taken as %E for linear hardening material. The FE analysis is carried out for different materials, that is, 500\u2264E/Y\u22651750. The stress values with respect to the tangent modulus of different materials are given in Table 2. The rigid sphere and a deformable flat contact model as shown in Figure 4. The load is applied on the top of the sphere. The sphere is penetrated into the flat. In this study an attempt has been made to modify the indentation depth in the new form by incorporating the tangent modulus, in terms of %E. The loading relationship for the penetration depth is given by the relation \u03c9={9L2/8D}1/3[ 2{(1 \u2013 \u03bd2) / (E* + ET) }]2/3 (1) In Eq. (1), L is the applied load, D is the ball diameter, and the paired material constants \u03bd, E*, and ET are the Poisson\u2019s ratio, equivalent Young\u2019s modulus, and tangent modulus, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002137_978-1-4614-1150-5_12-Figure12.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002137_978-1-4614-1150-5_12-Figure12.3-1.png",
        "caption": "Fig. 12.3 Internal forces and moments",
        "texts": [
            "1007/978-1-4614-1150-5_12, # Springer Science+Business Media, LLC 2012 165 If one of these two parts is considered, then the equilibrium condition requires that there is a force vector and/or a moment vector acting on the cut section to counterbalance the effects of the external forces and moments applied on that part. These are called the internal force and internal moment vectors. Of course, the same argument is true for the other part of the object. Furthermore, for the overall equilibrium of the object, the force vectors and moment vectors on either surface of the cut section must have equal magnitudes and opposite directions (Fig. 12.2). For a three-dimensional object, the internal forces and moments can be resolved into their components along three mutually perpendicular directions, as illustrated in Fig. 12.3. The force and moment vector components measured at the cut sections take special names reflecting their orientation and effects on the cut sections. Assuming that x is the direction normal (perpendicular) to the cut section, the force component Px in Fig. 12.3 is called the axial or normal force, and it is a measure of the pulling or pushing action of the externally applied forces in a direction perpendicular to the cut section. It is called a tensile force if it has a pulling action trying to elongate the part, or a compressive force if it has a pushing action tending to shorten the part. The force components Py and Pz are called shear forces, and they are measures of resistance to the sliding action of one cut section over the other. Their subscripts indicate their lines of action"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.13-1.png",
        "caption": "Fig. 7.13 A planar RR-type arm and its C-manifold as a flatted torus",
        "texts": [
            " Therefore, taking a 2D picture for the 2-torus T 2 will damage its topology, while taking a 3D stereo picture for T 2 can preserve all of its topological properties, such as its two \u201choles\u201d at a homotopical point of view. The multi-configuration issue will be very important for our next discussion on robotic dynamics and adaptive control, and the above example is just an explicit interpretation at both the geometrical and topological standpoints. Unfortunately, for an RR-type 2-link planar arm, or a robot having two revolute joints, whose axes are parallel to each other, its C-manifold becomes a flatted torus, as shown in Figure 7.13. This is also similar to the above case of squeezing T 2 into the 2D plane, resulting in an annulus in Figure 7.12. The z3 component of the original T 2 is now pushed down into the 2D plane to impose an additional projection on each of the first two components z1 and z2. According to Figure 7.13 on the right at point P , those additional projections are \u2212z3 sin \u03b81 = \u2212r sin \u03b82 sin \u03b81 for z1, and z3 cos \u03b81 = r sin \u03b82 cos \u03b81 for z2 such that the new equation for the flatted torus turns out to be { z1 = (a+ r cos \u03b82) cos \u03b81 \u2212 r sin \u03b82 sin \u03b81 = a cos \u03b81 + r cos(\u03b81 + \u03b82) z2 = (a+ r cos \u03b82) sin \u03b81 + r sin \u03b82 cos \u03b81 = a sin \u03b81 + r sin(\u03b81 + \u03b82). (7.44) This new equation coincidentally looks like the position vector p20 of the RR-type 2-link planar robot in symbolical form. The Jacobian matrix J can be immediately calculated as J = \u2202\u03b6 \u2202q = (\u2212as1 \u2212 rs12 \u2212rs12 ac1 + rc12 rc12 ) , where si = sin \u03b8i and ci = cos \u03b8i for i = 1, 2, and s12 = sin(\u03b81 + \u03b82) and c12 = cos(\u03b81 + \u03b82)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003871_itsc.2011.6082988-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003871_itsc.2011.6082988-Figure12-1.png",
        "caption": "Fig. 12. Two robots planning a rendezvous to a shared path",
        "texts": [
            " These sensors can also be simulated using building positions for LIDAR. Infrastructure is provided by working traffic lights and docking stations which can broadcast Vehicle to Infrastructure (V2I) messages. Player/Stage [8][9] is used to simulate sensors and propagation. For simulation, we use Player/Stage with multiple custom plugin libraries for matching the laboratory\u2019s setup. Figure 11 shows a top view of the simulation laboratory environment. Stage allows for 2.5D simulations, a perspective view is shown in Figure 12. 1) Radio Propagation: One of the custom plugins, calculates the radio propagation for the room\u2019s 802.11p\u2019s 5.85 GHz frequencies. Since the lab is a 1/10 scale model, first all distances are expanded by 10. The tagged mobile buildings positions and static structures are used to calculate line of sight, diffraction and reflection based propagation from each vehicle\u2019s position. This information is passed to the vehicles, so that potential receivers can be notified to not use a message if the power level is too low (<95dBm)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002503_isse.2011.6053543-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002503_isse.2011.6053543-Figure9-1.png",
        "caption": "Fig. 9. Model of substrate with thin layers of DPP2 and polystyrene protective layer.",
        "texts": [],
        "surrounding_texts": [
            "Two basic climatic cycles were used for finding out information about their electrical parameters response to ambient humidity and temperature. 4.1. Standard humidity-temperature test The thin layers were exposed to climate cycle in a chamber (Fig. 4). The temperature is changed in the range from 20 \u00b0C to 50 \u00b0C and relative humidity from 20 % to 90 %. The impedance of these layers were measured by precision LCR meter using four wires method at frequency of 1 kHz and voltage of 1 V. This test cycle was used in order to get the best response to humidity. 4.2. Extreme humidity-temperature test DPPs were exposed to the very rapid changes of relative humidity and temperature. The relative humidity was changed from 30 % to 85 % simultaneously with temperature from 30 \u00b0C to 85 \u00b0C (Fig. 5). These rapid changes were repeated twelve times. This test cycle was used in order to get the long-term stability of electrical parameters. The impedance of these layers were measured by using four wires method by means of precision LCR meter at frequency of 1 kHz and voltage of 1 V like in previous test."
        ]
    },
    {
        "image_filename": "designv11_100_0002706_20130825-4-us-2038.00035-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002706_20130825-4-us-2038.00035-Figure1-1.png",
        "caption": "Fig. 1.: Basic sensor with triple plate design",
        "texts": [
            " Monitoring the damage process during normal operation, below the tolerance limits of the components, enables specific preventive maintenance independent of rigid inspection intervals. 2. PRINCIPLES OF BASIC SENSOR With the oil sensor system, components of the complex impedances X of oils, in particular the specific electrical conductivity  and the relative permittivity r as well as the oil temperature T are measured [Gegner et al. 2008, Kuipers, Mauntz 2009]. The values  and r are determined independently of each other. Figure 1 shows the basic sensor with its triple plate design. Oils are electrical non-conductors. Thus the electrical conductivity of pure oils lies in the range below one picosiemens per metre (1 pS/m). For comparison, the electrical conductivity of another electrical non-conductor, distilled water, is larger by six orders of magnitude. Abrasive (metallic) wear, ions, broken oil molecules, acids, oil soaps, etc., all cause an increase of the oil conductivity . It rises with the increasing ion concentration and mobility"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.64-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.64-1.png",
        "caption": "Figure 2.64. Bypass circuits for induced negative field current: (a) crowbar; (b) varistor.",
        "texts": [
            " Under some conditions\u2014of pole slipping and system short circuits\u2014a negative current may be induced in the field of the synchronous machine. If this current is not allowed to flow, a dangerously high voltage can result across the field circuit. In some cases, damper windings or solid iron rotor effects may limit the maximum voltage experienced by the field winding and rectifiers under such conditions. In other cases, special circuitry is usually provided to bypass the exciter to allow negative field current to flow. These take the form of either \u201ccrowbar\u201d circuits (field shorting) or nonlinear resistors (varistors) (Figure 2.64a and b) [1,36]. In the case of the crowbar, a field discharge resistor (FDR) is inserted across the field of the synchronous machine by thyristors that are triggered on the overvoltage produced when the field current attempts to reverse and is blocked by the rectifiers on the output of the exciter. 6 Reprinted with permission from Ref. 36. (Reprinted with permission from Ref. 36.) Varistors are nonlinear resistors that are connected permanently across the field of the synchronous machine. During normal conditions, the resistance of these devices is very high, and little current flows through them"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003485_s1068798x12060093-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003485_s1068798x12060093-Figure2-1.png",
        "caption": "Fig. 2. Imbalance distribution over the generator rotor (a) and turbocompressor rotor (b).",
        "texts": [
            " Since information regarding the distribution of the residual imbalance over the length of the rotor is not available at the design stage (this distribution may be random and different even for rotors of the same type) and the imbalance may increase in the course of oper ation, the direction of the forces due to the imbalance must be selected in modeling on the basis of maximum increase in vibration and maximum flexure, dynamic stress, and reaction forces at each critical and specified working rotor speed. Since the distribution of the residual imbalance determines the degree of excitation of different intrinsic modes of rotor vibration, the most conservative approach for a flexible rotor is alignment of the imbalance vectors in sections of the rotor, in accordance with the intrinsic forms of rotor flexure. That case corresponds to maximum ampli tude of the transverse rotor vibrations in rotation. In Fig. 2, we show the calculation results for an imbal ance distribution corresponding to the third intrinsic form of rotor flexure observed at the frequency closest to the rated rotor speed. The damping characteristics of the system are determined by the internal friction (drag) in the rotor for steel with an absorption coefficient \u03c8 = 0.01\u20130.18, where \u03c8 is the ratio of the energy scattered within a single cycle to the energy at the beginning of the cycle [6]. The absorption coefficient is related to the relative damping coefficient: \u03be = \u03c8/4\u03c0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003497_iccis.2013.301-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003497_iccis.2013.301-Figure1-1.png",
        "caption": "Figure 1. The model of the full-vehicle dynamics system [1]",
        "texts": [
            " The parameters of current vehicle status are obtained in real time by using extended Kalman filter technique and the vehicle roll state is estimated based on LTR algorithm. The Matlab simulation shows this method is quite effective with easier process. II. THE EXTENDED 3-DOF NONLINEAR ROLLOVER The extended 3-DOF nonlinear rollover prediction model for heavy-duty vehicles is built based on principles of vehicle dynamics, and the relation between tire lateral force and tire side slip angle is presented by arctangent function. A. Rollover Prediction Model Based on Vehicle Dynamics Equation The model of the full-vehicle dynamics system is shown in Fig. 1, which is represented as a 3 DOF system. The lateral motion and two rotational (roll and yaw) motions are considered, longitudinal and vertical motions are supposed to be unchanged. Also, the wheels are considered to be symmetrical to x-axis. After applying a force-balance analysis to the model in Fig. 1, the equations of motion are given as \u03c8 zrrff JlFlF =\u2212 . (1) rfxy FFmhVVm +=\u2212+ \u03d5\u03c8 )( . (2) 978-0-7695-5004-6/13 $26.00 \u00a9 2013 IEEE DOI 10.1109/ICCIS.2013.301 11354 \u03c6\u03c6\u03c6\u03c6\u03c6 ry Jckmghhcocma =\u2212\u2212+ sin . (3) where m is the gross mass of vehicle (GM), Jx is the rolling moment of GM around center of gravity (CG), and Jr is the rolling moment of GM around center of rotation. Jz is the yawing moment of GM, lf is the length between front of vehicle and CG of GM, and lr is the length between rear of vehicle and CG of GM"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003635_amm.246-247.89-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003635_amm.246-247.89-Figure2-1.png",
        "caption": "Fig. 2 The mechanical model of bocce throw backhand hit",
        "texts": [
            "Therefore, we can say that the backhand throw to hit the technology in order to throw the ball with the right speed and strength, and the factors that affect the actual results are divided into the shot before the factors and the factors after the shot, influence throw throwfactors explain strike effect shown in Figure 1. From Figure 1, we can see all the factors, the shot rate, the shot angle and height is the key factor, they directly affect the effect of throwing strike. In order to study these three key factors is how it works, bocce throw backhand hit the mechanical model and mechanical analysis, bocce throw backhand hit the mechanical model shown in Figure 2. denoted by 0v . To the big ball in the process of pitching arm by the force, the arm force is denoted byF , the angle of the ball recorded as the ball height is denoted by\u03b1 , athletes shoulder 2H to ground distance in mind 1H , so you can get a backhand throw hit The mechanical model of mathematical expressions as follows: 2 1( ) sinH H L\u03b1 \u03b1= + (1) 0 1( )F F v\u03b1 \u03b1= \u2212 (2) Among them, L is for the horizontal component of the athlete's arm length, 0F for force, 1v for pitching arm after the natural rate of decline"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001940_978-3-642-21747-0_78-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001940_978-3-642-21747-0_78-Figure1-1.png",
        "caption": "Fig. 1. PMLSM drive system structure",
        "texts": [
            " This paper build bond graph model of PMLSM drive system. Unlike previous works, motor driver and sensor quantitative error are concerned. Model verification and analysis are obtained basing 20-sim software. System structure is firstly described in section 2 and sub-models of each component are built separately in section 3. Based on these sub-models, global model of PMLSM drive system is obtained in section 4 and system\u2019s characters are also analyzed by 20-sim software in this section. A typical PMLSM drive system is as shown in Figure 1[5]. It contains PMLSM, sensor, motor driver, controller and mechanism. Slider and mechanism moving along the guide, sensor real time collects position data and transmit it to controller, controller send a control signal to motor driver according position and control algorithms, finally, motor driver gives the current signal to armature realize the motion control. In this paper, ironless PMLSM motors are appropriate choices for its lack of detent force. Controller is not contains in global model because of there may be different algorithm in different systems and applications"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001669_978-3-642-33832-8_1-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001669_978-3-642-33832-8_1-Figure2-1.png",
        "caption": "Fig. 2 Phase photogram of self-excited vibration. a Soft self-excited vibration. b Hard selfexcited vibration",
        "texts": [
            " On the other hand, if the friction coefficient is at negative slope section, the vibration energy becomes higher and higher, which means that phase trajectory diverges more and more. Due to the existence of structure constrains, the friction coefficient swings alternately between positive slope section and negative slope section, that is to say, the vibration system balances between consuming energy and obtaining energy. Naturally, the self-excited vibration is generated. Self-excited vibration usually consists of hard self-excited vibration and soft self-excited vibration. As shown in Fig. 2a, soft self-excited vibration is unstable. With any kind of exciting force, the vibration system will diverge to a stable state. It is obviously to understand the mechanism from the phase photogram that the phase trajectory diverges to the stable limit cycle S1 at any initial situation. For a hard self-excited vibration, only if the initial exciting force is large enough to push the phase trajectory to pass the unstable limit cycle S2 and to achieve the stable limit cycle S1; otherwise, it will converge to the stable point O2:; which is shown in Fig. 2b. By analyzing the polygonal wear of tire, it is concluded that the tread\u2019s vibration is a typical hard self-excited vibration. Lupker [11] pointed out that the average wear rate of rubber would increase nonlinearly with the augment of vibration acceleration. As shown in Fig. 3, with the increase of vehicle speed, the normalized wear rate gets higher gradually, and then become smaller after the vehicle speed of 100 km/h. In another word, for a tire\u2019s self-excited vibration at the same frequency, different vibration accelerations can generate different vibration amplitudes, which cause different normalized wear rates"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001234_978-3-030-40513-7_24-Figure23-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001234_978-3-030-40513-7_24-Figure23-1.png",
        "caption": "Fig. 23 Prototype used for numerical modeling of heat conduction in composite component",
        "texts": [],
        "surrounding_texts": [
            "The finite element method (FEM) is a persuasive tool used in numerical methods to calculate approximate solutions to mathematical problems so that it can simulate the responses of physical systems to various forms of excitation. For analyzing thermal behavior of abovementioned composition, a numerical tool known as finite element method (FEM) enabled simulation of physical system response into excitation forms (Agrawal and Satapathy 2015). However, authors had employed ABAQUS 6.12 to analyze the effective insulation of composite for coating application in 3D print heads."
        ]
    },
    {
        "image_filename": "designv11_100_0003396_amm.275-277.174-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003396_amm.275-277.174-Figure3-1.png",
        "caption": "Fig. 3 Meshes of roller and stepped shaft",
        "texts": [],
        "surrounding_texts": [
            "Boundary Conditions and FE Model. The geometry parameters of the crankshaft section not shown in Fig. 1 are list in Table 3. Because the geometry and bending loads are symmetrical, quarter model of crankshaft section is selected. The geometry model is divided into some blocks, and then hexahedron meshes are generated as shown in Fig. 6. In the fillet region of main journal and rod journal where stresses concentrate, local mesh refinement approach is used. Symmetry constraints are set in the symmetry plane of model, and displacements in vertical direction are restrained by weak spring. After that, static bending moment of 2400 Nm is applied to the model. All the pre process operations are completed in HyperMesh. Results and Discussion. Under cycle bending loads, the amplitude of normal strain component \u03b5\u03b8 in dangerous section ,which could be marked as \u03b5\u03b8b, is closely related to the life of crankshaft, so it is always used in local stress-strain approach. The distributions of \u03b5\u03b8b in different depth a along \u03b8 are shown in Fig. 7. The maximum value of \u03b5\u03b8b is 5296 \u00b5\u03b5 in \u03b8=45\u00b0 of rod journal fillet surface."
        ]
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.43-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.43-1.png",
        "caption": "Fig. 2.43 4RPPaPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology R\\P\\kPa\\\\Pa",
        "texts": [
            "21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-Figure7-1.png",
        "caption": "Figure 7. Typical Simplified Compound Planetary GRA",
        "texts": [
            " Compared to the previous center-hinge type, the output capacity is almost cut in half because a single ring gear is used as the output. The other ring gear is used only for balance. The benefit is the ease of flange mounting and a smaller diametral envelope. We can simplify the compound planet gears to straight-cut planet gears that still mesh with two ring gears with carriers holding all the planet gears. A degenerate version of the previous cantilevered compound planetary GRA called simplified compound planetary GRA is shown in Figure 7, and the schematic of the compound stage is shown in Figure 8. Because two different ring gears mesh with the same planet gear, they are highly modified and limited by the number of planet gears that can be assembled. If sufficient envelope is available on the leading edge, the cost saving of simplified compound planetary GRAs could be beneficial. Another option is to have two different gears (non-straight-cut) on a compound planet instead of a straight-cut plant gear. This arrangement has the same advantages and disadvantages of the previous arrangement except with a larger diameter and a shorter length"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002859_amr.785-786.1172-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002859_amr.785-786.1172-Figure3-1.png",
        "caption": "Fig. 3 Stress nephogram(0\u00b0working condition)",
        "texts": [
            "6 steel balls are imposed on the line of fixed restriction. The contact stress is calculated through Equivalent Stress (Equivalent stress analysis module) ,and attain the stress nephogram that universal joint under 0\u00b0 working condition and limiting angle working condition when the car hangs one gear to start . Make the calculation of connecting stress of universal joint ,attain the stress nephogram that universal joint under 0\u00b0working condition and limiting working condition when the car hangs one gear to start . As shown in fig. 3, when the angle between the input and output shaft is 0\u00b0, the maximum stress is 2002.4 Mpa. The maximum stress of bell-like outer race is 977.56 Mpa, the maximum stress location is at Y \u2013axis 330\u00b0on outer race ; The maximum stress of steel ball is 2002.4 Mpa, the maximum stress location is at Y-axis 180\u00b0on the steel ball, it also is the weakest working condition ;The maximum stress of star-like inner race appears on inner race, the maximum stress is 1576.2 Mpa, the maximum stress location is at Y-axis 180\u00b0on inner race"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.15-1.png",
        "caption": "Figure 4.15 Large welded gear",
        "texts": [
            " Forged wheels are made solid or cored with round holes (Figure 4.14). The cored type is lighter, but requires more machining. The type without holes is simpler to work but is very heavy if of large width, and it does not allow homogenous mechanical properties to be obtained in the teeth after heat treatment. For a more convenient clamping of wheels on the machine tool, the web of the disc should be drilled between the rim and the hub. Sometimes large diameter holes are drilled to reduce the weight of the wheels. In a welded design of a wheel (Figure 4.15) the rim is connected to hub by two (or one) hollow discs welded to both rim and hub by a fillet weld. This connection is stiffened three times in six axial ribs. The hub is made of a rolled steel bar and the rim made of a steel sheet by bending. Gears of the largest diameter are often made in two parts. The plane of division passes through the middle of two ribs and two teeth spaces, and they are connected (with screws or welds) near the hub and near the rim. The thickness of the rim is usually three or four modules, and in carburized and hardened teeth it may be over six modules"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003531_mace.2011.5988269-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003531_mace.2011.5988269-Figure8-1.png",
        "caption": "Figure 8. Criterion of cracking.",
        "texts": [],
        "surrounding_texts": [
            "III. 4 720W 1mm/s\n40 38\n40 38 43.5 1400\n43.5 1400\n1Cr18Ni9 10\n5 7 x y z\n1400\n1400",
            "5 8\nZ X Y\n[11] 8\n1 3 2 Critical Strain rate for\ntemperature drop, CST\nCST\nCST",
            "9\n10\n- -\nIV.\n3082007\nREFERENCES [1] L. Dubourg and J. Archambeault, \u201cTechnological and scientific\nlandscape of laser cladding process in 2007,\u201d Surface & Coatings Technology, vol. 202, pp. 5863\u20135869, 2008. [2] F. Lusqui\u00f1os, R. Comesa\u00f1a, A. Riveiro, F. Quintero, and J. Pou, \u201cFibre laser micro-cladding of Co-based alloys on stainless steel,\u201d Surface & Coatings Technology, vol. 203, pp. 1933\u20131940,2009. [3] S. Barnes, N. Timms, B. Bryden, and I. Pashby, \u201cHigh power diode laser cladding,\u201d Journal of Materials Processing Technology, vol. 138, 411\u2013 416,2003. [4] L. Sextona, S. Lavina, G. Byrnea, and A. Kennedy, \u201cLaser cladding of aerospace materials,\u201d Journal of Materials Processing Technology, vol. 122, pp. 63\u201368, 2002. [5] L. Shepeleva, B. Medres, W. Kaplan, M. Bamberger, and A. Weisheit, \u201cLaser cladding of turbine blades,\u201d Surface and Coatings Technology, vol. 125, pp. 45\u201348, 2000. [6] A. Pinkerton and L. Li, \u201cMultiple-layer cladding of stainless steel using a high-powered diode laser: an experimental investigation of the process characteristics and material properties,\u201d Thin Solid Films, vol. 453, pp. 471\u2013476, 2004. [7] R. Jendrzejewski, G. liwi ski, M. Krawczuk, and W. Ostachowicz, \u201cTemperature and stress during laser cladding of double-layer coatings,\u201d Surface & Coatings Technology, vol. 201, pp. 3328\u20133334, 2006. [8] G. Zhao, C. Cho, and J. Kim, \u201cApplication of 3-D finite element method using Lagrangian formulation to dilution control in laser cladding process,\u201d International Journal of Mechanical Sciences, vol. 45, pp. 777\u2013796, 2003. [9] J. Kim and Y. Peng, \u201cMelt pool shape and dilution of laser cladding with wire feeding,\u201d Journal of Materials Processing Technology, vol. 104, pp. 284-293,2000. [10] Y. Jia and N. Hao, Thermal-mechanical Coupling Finite Element Analysis of Laser Cladding Process Part I: Temperature Field, Proc. ICEICE2011, in press. [11] T. Zuo, Laser processing of high strength aluminum alloys. Beijing, Defence Industry publisher, 2002 (In Chinese)."
        ]
    },
    {
        "image_filename": "designv11_100_0002485_s1068366613060123-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002485_s1068366613060123-Figure2-1.png",
        "caption": "Fig. 2. Photo of tribometer.",
        "texts": [
            " 1b), the drive moment is equilibrated by the moment of resistance Expression (4) also follows from this. The application of expression (2) to the boundary case of equilibrium on the inclined plane (Fig. 1b) yields the following relation: (5) or (6) where \u03b1 is the angle of inclination of the plane and \u03bcR is the static coefficient of friction. 0;xF =\u2211 0,yF =\u2211 0.AM =\u2211 ,e eF G N r r = \u22c5 = \u22c5 RM F r= \u22c5 .M N e\u00b5 = \u22c5 sin cos 0,G r G e\u03b1 \u22c5 \u2212 \u03b1 \u22c5 = tan ,R e r = \u03b1 = \u03bc Experiments were carried out using a tribometer with an inclined plane (Fig. 2) [1]. In this tribometer, contact pairs, along with a heater and a temperature sensor that are located very near to the zone of con tact, can rotate at a required angle \u03b1 with regard to the horizontal plane. The angle of inclination of the rotat able plane is mechanically implemented with an accu racy of up to 1'. The measuring system is isolated from the zone of heating. These experiments were aimed at determining the effect of the temperature on the static coefficient of friction under the rolling of balls with various radii over grooves with various radii of the generatrix"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.37-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.37-1.png",
        "caption": "Fig. 2.37 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PCPa (a) and 4PPaC (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology P\\C\\Pa (a) and P||Pa\\C (b)",
        "texts": [
            "33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.34-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.34-1.png",
        "caption": "Figure 12.34 A rolling disc on an inclined plane.",
        "texts": [
            "531) The B-expression of the Euler equation is usually the best option to analyze the rotational motion of a rigid body. The decoupling process of translation and rotation conceptually simplifies the analysis of motion of a rigid body. However, the equations of motion for translational and rotational motions are mathematically coupled and should be solved together. Example 752 Rolling Disc on an Inclined Plane A homogeneous disc with mass m and radius R rolls without slipping on an inclined ground plane such that the plane of the disc remains perpendicular to the ground. Figure 12.34 illustrates the rolling disc and the required coordinate frames. The ground plane is inclined by an angle \u03b1 with respect to the horizontal plane. The global frame G(I\u0302 , J\u0302 , K\u0302) is defined with gravitational acceleration g = \u2212gK\u0302, as shown in the figure. To indicate the inclined ground, we attach a temporary coordinate frame G1 (OX1Y1Z1) to the ground such that the Z1-axis is set perpendicular to the inclined ground. The ground plane is indicated by axes X1 and Y1. Let us attach a principal coordinate frame B(\u0131\u0302, j\u0302 , k\u0302) to the mass center C of the disc such that the y-axis is perpendicular to the disc plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003846_s1068798x11030129-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003846_s1068798x11030129-Figure3-1.png",
        "caption": "Fig. 3. Construction of the transition curve at the base of the gear teeth.",
        "texts": [
            " Given that z1 = z2, we obtain (2) x' xy \u03b8 \u03d5+( )sin yy \u03b8 \u03d5+( );cos+= y ' xy \u03b8 \u03d5+( )cos\u2013 yy \u03b8 \u03d5+( ).sin+= \u23ad \u23ac \u23ab P1' P2' xA aW \u03b32sin y ' \u03b32 z1 z2+( ) z1 sin x ' \u03b32 z1 z2+( ) z1 cos ;+ += yA aW \u03b32cos y ' \u03b32 z2 z1+( ) z1 cos x' \u03b32 z1 z2+( ) z1 sin .\u2013+= \u23ad \u23aa \u23aa \u23ac \u23aa \u23aa \u23ab xA aW \u03b32sin y ' 2\u03b32sin x ' 2\u03b32;cos+ += yA aW \u03b32cos y ' 2cos \u03b32 x ' 2sin \u03b32.\u2013+= \u23ad \u23ac \u23ab Taking values of \u03b32 from \u201318\u00b0 to 30\u00b0, we perform calculations on the basis of Eq. (2). From the coordinates obtained, we plot a loop shaped roulette, which is the trajectory of point A (the tooth tip of gear 1), in the system xO2y (Fig. 3). It engages with the involute curves of both tooth profiles without discontinuity and therefore may be adopted as the transition curve in formulating the gear tooth pro files. However, in gear rotation, the tooth tips will slip over the transition curves at the tooth bases; in other words, there will be no radial gap. Such outlines of the tooth bases are the outcome of the first stage of design. In the second stage, we must construct the radial gap required to compensate the imperfections in gear manufacture and installation, lubricant distribution, etc. In the transmission, in accordance with the tooth depth, a maximum radial gap of 5 mm is assumed. However, its outline must be consistent with simple manufacturing technology and minimum stress con centration at the tooth root. As is evident from Fig. 3, the radial gap contour in the present case includes arcs of circles and straight line segments. Other tran sition curves are possible\u2014for example, a single arc covering the whole of the roulette. If the transmissions have different numbers of gear teeth and pinion teeth, these design stages must be repeated appropriately. (1) In designing the transition curve at the tooth base in an involute transmission with asymmetric tooth profiles, the first step is to calculate the prelimi nary curve, with no radial gap between the conjugate gear teeth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003321_s12206-011-0630-6-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003321_s12206-011-0630-6-Figure2-1.png",
        "caption": "Fig. 2. Free body diagram at node (i,j) in a string bed.",
        "texts": [
            " Results showed that smaller tension allows more power and less impact on the arm, while larger tension offers more control, and that the resultant force on the player\u2019s hand is larger if the tennis ball hits the dead spot than if it hits the sweet spot. The objective of this study was to investigate the structural behavior of the string bed of a tennis racket subjected to transverse force perpendicular to the string bed plane. The mathe- matical formulation derived for the string bed was implemented into a computer programming code, which was used to conduct parametric studies on the structural behavior of the string bed. Fig. 2 depicts the segment cut out from a typical string bed. The dotted lines in the figure represent the deformed configuration of the string bed. While the x-z plane represents the plane of the undeformed string bed, strings running in the directions of z and x are defined as main string and cross string, respectively. The main string and the cross string perpendicular to each other are assumed to possess linear elasticity and have no friction at their intersection. External transverse force perpendicular to the x-z plane acts on a node, an interaction of the main and cross strings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003009_imece2013-62977-Figure14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003009_imece2013-62977-Figure14-1.png",
        "caption": "Fig. 14: Catapult that utilizes an elastic band or spring [10]",
        "texts": [
            " Downloaded From: https://proceedings.asmedigitalcollection.asme.org on 04/09/2019 Terms of Use: http://www.asme.org/about-asme/terms-of-use 8 Copyright \u00a9 2013 by ASME After the validation of the design requirements and much research the MAE design team narrowed down the design concepts to three options. Two of the concepts were very similar whereby they utilized the elastic force consisting of a spring or an elastic band to transfer energy to the projectile. An example of this type of design is shown in fig. 14. The projectile would be placed in the projectile cavity located on the end of the lever arm. As the operator presses down on the lever arm, potential energy is stored in the spring or elastic band. When the lever arm is released the potential energy is converted to kinetic energy that enables the launch of the projectile. The MAE design team considered this concept as it would provide a simple yet effective manner to launch the projectile [9]. The third design concept considered by the MAE design team [9] was a trebuchet which utilizes a counter-weight to provide the kinetic energy to the projectile that is positioned within a sling-as shown in fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure7-1.png",
        "caption": "Figure 7. Input and Output Force",
        "texts": [],
        "surrounding_texts": [
            "No muscle is attached to clavicle. Therefore, the articulation sternoclavicularis and the acromial joint are passive joints, and only the articulation acromioclavicularis and the scapulothoracic joint are active ones. The scapula is attached to the sternum via the clavicle. Therefore the scapula motion is restricted to the spherical surface whose radius is the clavicle. Skin also restricts the scapula motion so that the scapula is always on the ribs.\nFig.3(a) shows a schematic model of these relations, and Fig.3(b) presents our 4DoF scapula model. In this model, the scapula is approximated as a rectangular board, the clavicle and the rib are approximated as links which have 3DoF joints at both ends. The 4DoF model has four sliding actuators.\nFig. 4(a) is the back view of the human body. Red areas are soft areas to which large force cannot be applied from outside. White areas in Fig. 4(a) are hard areas to which outside force can be applied. The only hard areas around the shoulder are around the articulatio acromioclavicularis and around the scapulothoracic joint. These are the points to which the trainer will apply guidance force to let the trainee learn the proper motion of the scapula and the shoulder joints. Fig. 4(b) shows corresponding points in the 4DoF model.\nIII. SCAPULA MOTION AND SCAPULA MOTION ANALYSIS\nA. Aanlysis Method by the 4DoF Model\nIn Fig. 5, \u2211 0 is the coordinate system on the human body,\nand \u2211 p is the coordinate system on the scapula ( 4321 CCCC )\nwith its origin at p O . Position and pose of the scapula in the\n4DoF model is expressed by four angles ( 1 \u03b8 , 2 \u03b8 , 3 \u03b8 , 4 \u03b8 ). 1 R ,\n2 R , 3 R , 4 R represent length variables of four sliding actuators.\nThe angle 1 \u03b8 is the angle of 113 BAA . The angle 2 \u03b8 is the angle between the triangle 113 BAA and the XY plane of \u2211 0 . The angle 3 \u03b8 is the angle between the triangle 113 BAA and the\ntriangle 331 BAB . The angle 4 \u03b8 is the rotation angle of the quadrilateral 4321 CCCC whose rotational axis is 31 BB XY . 1 a ,\n2 a , 1 b , 2 b , 1 l (length of 11 BA ) and 2 l (length of 33 BA ) are constants which are obtained by measuring the human body. 1 \u03b8\ncorresponds to elevation and depression of the scapula. 2 \u03b8 corresponds to forward and backward motion. 3 \u03b8 and angle 4 \u03b8 correspond to rotation of the scapula.\nThe transformation matrix p\nT0 between the coordinate\nsystems \u2211 p and \u2211 0 is presented by 1 \u03b8 , 2 \u03b8 , 3 \u03b8 , 4 \u03b8 . Between\nthe coordinate of a vertex i C ( 1=i , 2 , 3 , 4 ) in \u2211 p and its\ncoordinate in \u2211 0 , the following relation holds:\ni\np pi CTC 00 = \uff081\uff09\nAs the transformation matrix p T0 is expressed by 1 \u03b8 , 2 \u03b8 ,\n3 \u03b8 , 4 \u03b8 , the equation (1) can also be described as:\n( ) i p i CgC P=0 \uff082\uff09\nwhere [ ]T 4321 \u03b8\u03b8\u03b8\u03b8=P .\nWith the coordinate of the origin p O of \u2211 p in \u2211 0 ,\n[ ]T pzpypxp OOOO 0000 = , and the pose of the scapula roll \u03b8 , pitch \u03b8 ,\nyaw \u03b8 , the position and pose of the scapula in \u2211 0 can be described as [ ]T yawpitchrollpzpypx OOO \u03b8\u03b8\u03b8000=Q , and\nanother expression of the equation (1) is obtained as:\n( ) i p i ChC Q=0 \uff083\uff09\nTherefore,",
            "The force from the other actuators can be expressed in the same fashion. Then, the force balance in the 4DoF is represented by the following relations.\n( ) ( ) ( )\n \n  \n\n\u2211=\n\u2211=\n\u2211=\n=\n=\n=\n6\n1\n00\n6\n1\n00\n6\n1\n00\ncos\ncos\ncos\ni iizz\ni iiyy\ni iixx\nFF\nFF\nFF\n\u03b3\n\u03b2\n\u03b1\n\uff088\uff09\nConsidering the center of the gravity of the scapula\n[ ]T gzgygxg OOOO 0000 = , and the point [ ]T azayaxa OOOO 0000 =\nwhere the outside force F is applied , the torque balance of the 4DoF model is represented as follows.\n( ) ( ) ( ) ( )\n( ) ( )\n  \n\n    \n\n\u2211         \u2212 \u2212 =\n\u2211         \u2212 \u2212 =\n\u2211         \u2212 \u2212 =\n=\n=\n=\n6\n1\n0\n0\n6\n1\n0\n0\n6\n1\n0\n0\ncos\ncos\ncos\ni gzaz\niigziz\nz\ni gyay\niigyiy\ny\ni gxax\niigxix\nx\nOO\nFOC F\nOO\nFOC F\nOO\nFOC F\n\u03b3\n\u03b2\n\u03b1\n\uff089\uff09\nC. Scapula Motion Analysis and Force of Shoulder Mouscles\nAnalysis\nScapula motion and force of shoulder muscles are analyzed by the 4DoF model for elevation, depression, forward, and backward motions of the scapula. The force of the four actuators is regarded as the force from the muscles.\nIn this analysis, it is assumed that the human is standing upright in the natural posture, and the arm hangs from the",
            "shoulder without exercising any tasks. Therefore, the input force F to the scapula is the weight of the arm.\nConstants in the 4DoF model (Fig. 3) need to be measured in advance from the individual trainee. In this paper, average values of Japanese are used [7]. In order to calculate the center of the gravity in the 4DoF model, the weight ratio of various parts of the human bones is also necessary [7]. Table I and Table II summarize these constants, the input force F , and the weight ratios.\nFour actuators\u2019 force was calculated for a couple of simple scapula motions by the method described in the previous section. Range of scapula motion is determined from Range of Motion (ROM) [8]. When the man stands upright with his arm hanging naturally,\n1 \u03b8 , 2 \u03b8 , 3 \u03b8 , and 4 \u03b8 are 90 degrees, 90\ndegrees, 6 degrees, and 0 degree respectively. If the scapula is elevated or depressed,\n1 \u03b8 , 2 \u03b8 , 3 \u03b8 , and 4 \u03b8 change in the ranges\nshown in Table III. The scapula at 70 1 =\u03b8 degrees corresponds to the posture where the shoulder is moved to the uppermost position. When the shoulder is at the lowermost position,\n1 \u03b8 is\n110 degrees. Likewise, if the shoulder is moved from the backward limit position to the forward limit position,\n2 \u03b8\nchanges from 70 degrees through 110 degrees (Table IV).\nFig.8 shows scapula\u2019s position p O pose roll \u03b8 , pitch \u03b8 , yaw \u03b8 , and\nFig.9 shows the forces from the four actuators, when the shoulder is moved in the range shown in Table III. Fig.10 and 11 show the position, pose, and forces calculated for the forward and backward motion in Table IV."
        ]
    },
    {
        "image_filename": "designv11_100_0003079_amm.182-183.1529-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003079_amm.182-183.1529-Figure1-1.png",
        "caption": "Fig. 1 EHL of line contact and equivalent geometry figure",
        "texts": [
            " Because water has no pollution, sourced widely, cheap, safe, not burning and other characteristics, so it's a kind of lubrication medium which is very potential for the development. In the past the EHL research of water-lubricated rubber bearing has focused on the so-called state of the full water supply condition, the starved conditions of water-lubricated rubber bearing have not been analyzed. This paper will solve the problem. According to the line contact elastohydrodynamic lubrication theory, the EHL model and equivalent geometry model is shown in figure 1 below. 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: 142.103.160.110, University of British Columbia, Kelowna, Canada-12/07/15,08:49:44) Reynolds Equation The Reynolds equation is given by [2]: ( )3 12 e d hd h dp U dx dx dx \u03c1\u03c1 \u03b7   =    (1) Fig2 Schematic of oil provided parameters As shown in Fig. 2, it is assumed that both surfaces a and b bring a thin layer of lubricant into the contact, and the thicknesses of the oil layers are ah and bh , respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003213_1.4789220-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003213_1.4789220-Figure2-1.png",
        "caption": "FIGURE 2. Schematic diagram of the non-contacting thermoelectric technique.",
        "texts": [
            " The contact effects are related to the imperfect contact between the test sample and the reference probe, amount of pressure applied to the probe, temperature of hot and cold junctions and probe material [12]. However, thermoelectric measurement can be conducted in an entirely noncontacting way by using high-sensitivity magnetic sensors to detect the thermoelectric current caused by inclusions and other types of inhomogeneities when inspected specimen is subjected to direct heating and cooling as shown in Figure 2. Assuming the existence of a defect or imperfection in an otherwise homogeneous material and that a temperature gradient is established throughout the specimen, different points at the boundary between the defect or imperfection and the host material will be at different temperatures, and therefore at different thermoelectric potentials. These potential differences will drive local thermoelectric currents around the affected area, which can be detected in a noncontacting manner by a high sensitive magnetometer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure6-1.png",
        "caption": "Figure 6. Nurring Structure of Bushing Steel Pipe.",
        "texts": [
            " And to combine ball stud, upper and lower bearings were designed as a locking structure as shown in Figure 5. Suspension bush is like the joints for human. And the role of the suspension bushings is absorbed the forces caused ball joint. But because bushing outer pipe material is steel or aluminum, bushing and outer pipe occur to slip and each other is lack of pull-out strength. If it is covered plastics after nurring a steel pipe, steel pipes and plastics have a strong combination and pull-out strength degradation does not occur, as shown in Figure 6. If it is designed in shown a Figure 6, current forged upper arm of production is equal to pull-out strength. As shown in Figure 7, Ball joint pipe and bushing pipe were welded to center steel press and in order to increase in the durability and strength of welds, welds part were covered with plastics. Engineering plastics is materials that improved which thermal properties and mechanical strength. But even if it has a heat resistant, chemical resistant and high strength, engineering plastic is absolutely necessary considering of the safety factor properties because it cannot have permanent properties like steel and aluminum"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001945_978-1-4614-1788-0_13-Figure13.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001945_978-1-4614-1788-0_13-Figure13.1-1.png",
        "caption": "Fig. 13.1 ( a ) Schematic drawing of MQ-MD system. The slow fl ow into the hollow fi ber membrane generates a dialysate, which is rapidly diluted and removed by a second, fast \u201cdilution\u201d fl ow resulting in a \u201ctotal fl ow\u201d out of the probe. ( b ) An ultrafi ltration collection device (UCD) consists of a Monovette under-pressure container, a capillary collection tube and a hollow fi ber (next to the paperclip). ( c ) A schematic drawing of the UCD with enlargement of the hollow fi ber membrane, containing a spiral, which prevents the hollow fi ber from collapsing. The spiral extends into the restriction tube, which defi nes the fl ow rate in combination with the under-pressure container used",
        "texts": [
            " Another problem of using us-fl ow rates is the very long lag times between sampling and analysis. In addition, evaporation may occur during the long collection times, which concentrates the samples. These disadvantages can be circumvented by using the MetaQuant (MQ) sampling technique (patented by Brainlink B.V.), which allows fast transport of dialysate together with high recoveries of the analytes. In this type of MD probe an extra fl uid line is added, which rapidly transports the dialysate sample to the detection device (Cremers and Ebert 2007 ) (see Fig. 13.1a ). A drawback of the additional fl uid line is that extra care should be taken to keep the sampling system sterile. The driving force of the fl ow rate in MF is a pressure gradient over the probe membrane, pulling sample from the in vivo environment. MF samples are undiluted (100% recovery), small ( m l or nl size) and sterile. Figure 13.1c is showing a typical MF device schematically. Two different types of MF probes have been developed of which only one is commercially available (Bioanalytical Systems, Lafayette, Ind., USA). The commercial probe consists of one or more hollow fi ber loops with various hollow fi ber lengths (ranging from 2 to 12 cm). The sampling rate is relatively high, in the range of 0.5\u20132 m l/h/cm hollow fi ber membrane. A previously described non-commercial probe is much smaller: a concentric probe of 4 cm using a us-sampling rate of 100 nl/min, applied subcutaneously in conscious rats (Moscone et al . 1996 ) . In this study the MF probe was connected to a long capillary, enabling continuous sampling for 30 h and simultaneous storage of the samples in that collection coil. This integrated device is called the ultra-fi ltration collection device (UCD; Fig. 13.1b, c ). With this device time-profi les can be created without manual intervention and the profi le can be stored in de collection device until further analysis up to several days without signifi cant diffusion artifacts. The collection coil is made of fused silica tubing, because a signifi cant evaporation of the collected fl uid through several other tested materials was observed. Due to the small size of the probe, the probe can be introduced in vivo via adapted needles with minimal tissue damage and without using anesthetics"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003948_emeit.2011.6023972-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003948_emeit.2011.6023972-Figure1-1.png",
        "caption": "Figure 1. Expansion plan for meshing pair of oscillating-teeth end-face harmonic gear",
        "texts": [
            " Known from meshing theory of oscillating-teeth end-face harmonic gear, In order to facilitate the design and manufacture, in the meshing pair A and B, bus bar of teethface can be designed Archimedes spiral face which bus bar is line and vertical and cross to rotation middle line. When the wave generator rotates in counterclockwise constant angular velocity, oscillating teeth driven by the wave generator's rise surface begin to enter meshing state, and move from the tooth tip to tooth root of the end-face gear, at the same time, constrained by Vice-axle guide sheave, oscillating teeth push sheave clockwise rotation. When wave generator rotates to its highest point, oscillating teeth get to the location of oscillating teeth 1 as shown in figure 1 and contacts end-face gear's tooth bottom, and working stroke of meshing pair is going to stop, when the wave generator continues to rotate, meshing pair begins to enter empty run, under the action of the end-face gear 's anti-push , oscillating teeth return to original position along non-work teeth-faces of end-face gear, complete a motion cycle of meshing pair. All oscillating teeth repeat that mesh movement under the condition of keeping etc spacing. Because meshingzone and non-meshingzone rotate together with wave generator, six parallel meshing pairs of oscillating-teeth end-face harmonic gear orderly move in the 978-l-61284-088-8/ll/$26",
            " MESHING AREA CHANGING DISCIPLINARIAN OF OSCILLATING-TEETH END-FACE HARMONIC GEAR'S WORK MESHING PAIR Transmission types of oscillating-teeth end-face harmonic gear have two kinds which one is one sided drive; the other is double sided drive. One sided drive only has a set of meshing end-face gear, oscillating teeth and wave generator. But double sided drive has wave generator which have two gearing faces symmetrically distributes in the twoside surface of wave generator ,then the wave generator respectively meshes the two sets of oscillating teeth and endface gears. sided drive's work meshing pair Figure 1 shows the instantaneous meshing state of each oscillating teeth, oscillating teethl,5 and 6 engage in the work meshing condition, oscillating teeth 2 and 3 engage in the non-work meshing condition, and oscillating teeth 4 is undocked. Because oscillating teeth 1 fully meshes end-face gear, total meshing area of work meshing pairs reaches maximum. In the next instant, oscillating teeth 1 will quit work meshing and enter non-work meshing, and oscillating teeth 4 will enter work meshing, because meshing area of oscillating teeth 1 suddenly reduced to zero, and meshing area of other work meshing pair can be regarded as constant, total meshing area of work meshing pairs reaches minimum [5], it can be seen from that, figure 1 shows a mutation moment when instantaneous value of total meshing area changes from maximum to minimum. The change process for total meshing area shows following :The moment is calculated when oscillating teeth 1 is just undocked, at the moment, oscillating teeth 4 engage in the work meshing condition, next oscillating teeth 4,5 and 6 engage in the work meshing condition, and the meshing area of three oscillating teeth will gradually increase. So the total meshing area will gradually increase in this process until oscillating teeth 6 reach the fully meshed state, at this time the total meshing area achieve maximum"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003495_amm.226-228.1665-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003495_amm.226-228.1665-Figure6-1.png",
        "caption": "Fig. 6 Mechanism for Climbing Up and Down",
        "texts": [
            " The two sets of suction cups move in relation to each other to create rotational motions by one rotary actuator and linear motions by two linear actuators. The power is supplied through electrical cables from ground. Figure 5 shows the CCD camera and the transmitter installed in the remote-controlled visual inspection machine. The video signal from the CCD camera is converted into an analogue television signal by a transmitter. The analogue television signal is converted back into visual data. The signal can be received within a maximum distance of 50 m to 150 m depending on conditions. Figure 6 shows the configuration of the machine to ascend. Figure 6 (a) shows the configuration of the machine when the suction cups A and C are in operation to support the main body of the machine. In the next stage in Fig. 6 (b), the suction cups B and D are lifted up. The suction cups B and D are in operation, and the suction cups A and C are out of operation as shown in Fig. 6 (c). Finally, the main body of the machine is lifted up by the two linear actuators. Repeating these operations, the climbing testing machine is able to climb up walls. The inverse operation lets the climbing testing machine climb down walls. Four suction cups are always in operation at one time. One suction cup can hold 100 N. Therefore, this machine can theoretically hold more than 40 kg in weight. However, real concrete walls may be uneven or have grooves. When suction cups are adhering to uneven or grooved parts of concrete walls, the maximum holding force may decrease"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.46-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.46-1.png",
        "caption": "Fig. 2.46 4PaRPPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa||R\\P\\\\Pa",
        "texts": [
            "22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig. 2.49) R\\Pa||Pa\\kR (Fig. 2.22j) The first and the last revolute joints of the four limbs are parallel 38. 4RPaPaR (Fig. 2.50) R\\Pa||Pa\\kR (Fig. 2.22k) Idem No. 37 39. 4RPaRPa (Fig. 2.51) R\\Pa\\kR\\Pa (Fig. 2.22l) Idem No. 15 40. 4PaRPPa (Fig. 2.52) Pa\\R\\P||Pa (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003298_amr.179-180.150-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003298_amr.179-180.150-Figure1-1.png",
        "caption": "Figure 1: Yarn feed mechanism of carpet tufting machine",
        "texts": [
            " The cam drive the needle bar transversely shifting the relative position in respond to predetermined stitch pattern information in cam profile. In multi-pile-height jacquard, the gap in tufting surface has fatal affects in carpet quality. Therefore, it is necessary to accurately adjust the yarn delivery and eliminate the additional pile height change caused by needle bar transverse. The work reported in this paper is concerned with yarn feeding compensation solution by dynamic control yarn feeding in each tuft cycle to precise adjust yarn delivery amount in the manufacture of loop-pile tufted carpet. Fig.1 is a schematic diagram illustrating the typical yarn feed mechanism of carpet tufting machine assembled with jacquard mechanism of clutches. A tuft cycle start from top dead center (TDC), the needle threaded with yarn is pushed down through a backing fabric(cloth)[8-12]. As the needle approaches bottom dead center (BDC), a looper advances and passes between the needle and the yarn. The needle retracts, to leave a loop of yarn around the looper. This loop of yarn is released intact to produce a loop-pile fabric",
            " The needle bar transverse will affect both in needle tracking and yarn guide route, which generates additional yarn consumption represented with L\u2206 . The additional usage of yarn length due to varying needle tracking is marked as tL\u2206 and the additional length due to varying yarn guide route is marked as yL\u2206 , and get L\u2206 as Eq. 1. t y= +L L L\u2206 \u2206 \u2206 (1) As Fig. 2 shown is the yarn usage in rectangular coordinate. Before feeding the yarn to individual needles, it needs to cross over two sets of thread-carriers from the yarn feed roller. One fixed guide plate as 10 in Fig.1 is fixed on the machine frame, which facilitates the guiding of the yarn onto the yarn feed roller. Another two guide plates are oscillatory with needle bar. The alternation of relative place for switching between the two thread-carriers is used to store yarn and adjust yarn tension. As shown in Fig. 2, consider the yarn as a rigid body and set up the rectangular coordinate, the grid origin is the fix outlet terminal of fix guide carrier. The positive y-axis is to direct the motion of the backing cloth"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002430_amm.457-458.581-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002430_amm.457-458.581-Figure1-1.png",
        "caption": "Fig 1 The inflatable column structure local compression model (a) 3D schematic model (b) 2D side view of the collapse zone",
        "texts": [
            " Quantitative index of the concept was given subsequently for inflatable column structures. Experimental and numerical methods were performed to obtain the shape retention index, through which we can make a comparison of shape retention quantitatively. The shape retention index of an inflatable column structure and its index The shape retention was defined as a global property. To make sure the load affected zone is local, the dimension of inflatable column must far larger(10 times in usual) than the dimension of the circle compression die(Fig 1(a)). 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: 142.103.160.110, University of British Columbia, Kelowna, Canada-13/07/15,18:04:27) When an inflatable column structure is compressed by a certain height , the compression shape changes significantly, so it is necessary to describe the compressive deformation first. Compressive deformation. A dimensionless index was introduced to depict the change of the compression shape. The 3D compression process was taken as a 2D plain problem for simplicity. The upper border of the inflatable column is a straight line from side view, and the circle die affected zone will collapse after compression. A trapezoid was used to represent the collapse region(Fig 1(b)). The area ratio was defined by area change, and the length ratio was defined by length change similarly. 2 1 _ S area ratio S  . 2_ tot l length ratio l  . (1) S2 is the area covered by the inclined line in Fig 1(b) , which is a preconceived region that will be affected. S1 is the trapezoidal which is actually affected after compression. L2 is the length of the straight line before deformation; ltot means the length of the rest of segments of the trapezoid. Compressive deformation shape index was defined as a sum of the area ratio and the length ratio. The shape index increases monotonously with the deformation angle \u03b8, even linearly in some extent(Fig 2). It seems that dimensionless compression shape index depicts the ability of shape retention of the column"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003758_omae2013-11433-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003758_omae2013-11433-Figure1-1.png",
        "caption": "FIGURE 1. DEFINITION OF THE EARTH-FIXED AND BODYFIXED FRAMES",
        "texts": [
            " The influence of the vessel draft, the relative incidence of the current, waves and wind, and the variation of the peak period of the sea wave spectrum are analyzed. All evaluations are performed by dynamic simulation using a nonlinear mathematical code. The mathematical model expresses DPS vessel motions in the horizontal plane under the action of the environmental and actuator forces and moment. It is derived from equations of kinematics and dynamics and is determined taking into account two references, as shown in Fig. 1. The first, OXY Z is an earth fixed frame, which can be considered an inertial frame. The second reference frame is a non-inertial body fixed frame GXGYGZG. In general, the position of the reference is such that its axes coincide with the principal axes of inertia of the vessel. As the main concern of the DPS is the study of the vessel horizontal motion it is convenient to adopt planes OXY and OXGYG as parallel to the sea surface. In Fig. 1 the direction of the current, wind and waves are indicated also. Additional conventional assumption is to split the total motions of the vessel in each direction as the sum of the low and wave frequencies motions. Besides, in this paper are considered the saturation of the actuators. Under those assumptions the lowfrequency motion of the vessel can be expressed from the following equations: \u03b7\u0307 = J(\u03c8)\u03bd (1) (MRB +MA)\u03bd\u0307 +[CRB(\u03bd)+CA(\u03bd)]\u03bd \u2212MA\u03bd\u0307c +CA(\u03bd)\u03bdc \u2212CA(\u03bdc)\u03bdr = F +B fP (2) Where: J(\u03c8) = cos(\u03c8) \u2212sin(\u03c8) 0 sin(\u03c8) cos(\u03c8) 0 0 0 1  (3) 2 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001688_9780857094537.3.135-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001688_9780857094537.3.135-Figure7-1.png",
        "caption": "Figure 7: a) Equally spaced lower balls,",
        "texts": [
            " as soon as a spall disrupts the smooth rolling contact of surfaces. Run-down from this trigger to complete halt takes no longer than 3 seconds, so it is reasonable to conclude that these AE signals show signs of increasing activity before significant surface damage has occurred. However, clearly the raceway was not contributing to the 2xBPFO modulating frequency. The initial assumptions for BPFI and BPFO in this research assume that the three lower balls are equally spaced, but closer inspection reveals that this may not be the case. Fig. 7 has been prepared using the exact geometry of the ball race and balls, demonstrating that space exists for the three balls to re-align themselves in proximity. The balls are numbered 1, 2 and 3 for reference. b) maximum possible re-alignment (approximate angles) Assuming that the three lower balls actually align themselves such that they nearly touch as in Figure 7b, the theoretical pulse wave which would be created from a stress cycle at BPFI was created (Figure 8). The pulse representing ball 2 is lower in amplitude, in common with observations from the actual AE data from the tests. As an aside, this may explain a commonly observed phenomenon of the four-ball machine relating to the condition of the lower balls after a test. Two of the three balls consistently emerge with clear rotational markings whereas the third (assumed here to be ball 2) emerges with no particular tracks and random scuffing over the entire surface (Figure 9)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003398_amm.229-231.406-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003398_amm.229-231.406-Figure1-1.png",
        "caption": "Figure 1 (b) shows contact trace of WN gear tooth surface. WN gear with involute gear difference is that meshing deputy campaign does not mobile along direction of tooth height but along the spiral surface of tooth width direction. In the meshing transmission process, contact point of WN gear in the position of the tooth direction process will move up and down with the changes of center distance.",
        "texts": [
            " In this study, based on the gear contact load changes, considering the systematic errors and the WN gear meshing characteristics, the contact of the gear drive and strength analysis model is established, which space geometry theory is used. Meshing characteristics and simulation analysis of WN gear Normal surface tooth profile curve of WN is drawn, which reference of GB/T 12759-1991 normal surface tooth profile curve of double circular cylindrical gear. Basic tooth profile of WN gear is shown in Figure 1 (a): It consists of four work arcs which are convex arcs, concave, connected arc of the convex and concave arc, tooth root arc. Each segment of the arc is determined by the radius and the center eccentric position. Detailed tooth profile parameters and data of WN gear can be found in the relevant design standards and information. The modeling process is as follows: Firstly, benchmark curve is created by tooth profile equation of WN gear. Secondly, the contour of normal surface tooth groove is established along the helix alveolar by scanning mixed-cutting feature",
            " Position of tooth contact and shape of contact area is clearly visible in the shadow. Changes in the results of contact mode are simulated by changes of the helix angle, center distance and the contact load. Compared with the theoretical analysis, meshing feature and strength properties of the WN gear driving are gained. The study found that the contacts region shown in the shaded when the helix angle is small, the long axis of oval with meshing path parallel. With the increase of helix angle, as shown in figure 1(b) the long axis of oval with meshing path vertical. Contact area and contact analysis model of WN gear According to the spatial relationship of the two contact surfaces, convex and concave tooth surface of the arc gear meshing can be seen approximately as surface contact of a football shape and a saddle-shaped body [4,5]. Contact of arc gear is theoretically the point, but under normal circumstances meshing contact area of WN gear is an ellipse or rectangle, which is caused by contact elastic deformation and convex and concave surface curvature changes after running"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.25-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.25-1.png",
        "caption": "Figure 3.25 The stability axes, (ie, ke), the displaced body axes (i, k), and the forces and moment for longitudinal motion",
        "texts": [
            " The flight variables in such an equilibrium are denoted with a subscript e: Ue, We, \u03b8e, Qe, Xe, Ze, Me. A common equilibrium point is an essentially straight-line flight path in the longitudinal plane, for which Qe = 0 and equation (3.145) yields \u03b8\u0307e = 0, Xe \u2212 mg sin \u03b8e = 0, Ze + mg cos \u03b8e = 0, Me = 0. (3.146) A convenient reference frame for longitudinal motion is the stability axes, (ie, je, ke), defined such that the axis ie is along the equilibrium flight direction,We = 0. Thus, a longitudinal displacement from equilibrium can be described by the displaced body-fixed frame, (i, je, k) (Figure 3.25), as well as perturbations in flight variables and aerodynamic forces and moment given by: = \u03b8e + \u03b8, U = Ue + u, W = w, Q = q, X = Xe + X\u0304, Z = Ze + Z\u0304, M = M\u0304. (3.147) In such a case, the vertical velocity perturbation, w (called downwash), is represented as a change in the angle of attack (Figure 3.25) given by \u03b1 = tan\u22121 w U . (3.148) The perturbation quantities, \u03b8, u, w, q, X\u0304, Z\u0304, M\u0304, can be considered small to begin with. Since a successful flight control system would prevent the small perturbations from growing large with time, one need not consider large aerodynamic perturbations. This is fortunate because aerodynamic forces and moments have a rather complex character in the case of large perturbations, which can be very difficult to analyze in closed form. A small perturbation from the equilibrium results in the following approximations: sin sin \u03b8e + \u03b8 cos \u03b8e, cos cos \u03b8e \u2212 \u03b8 sin \u03b8e, u Ue, w Ue, \u03b1 tan\u22121 w Ue w Ue , and thus the rotational kinematics equation (3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002094_978-94-007-4902-3_74-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002094_978-94-007-4902-3_74-Figure2-1.png",
        "caption": "Fig. 2 Analysis of Walshaert mechanism mobility by two groups (a) and (b): first (left) and second (right) attempts",
        "texts": [
            " When marking the second submissions the process by which students achieved this, in general, became clear. As they had to resubmit the first submission, it often contained their \u201cmarking\u201d. This assisted in recognizing the level of feedback/iteration within groups. An example of how a group digested tutor comments and recommendations of some \u2018good groups\u2019 members is shown in Fig. 1. Comparison between the first (left) and second (right) attempt at calculating the mobility of the 2D Walshaert mechanism is shown in Fig. 2. Figure 2a is the same group as Fig. 1. Figure 2b is by another group. In both cases, the hand written notes on the left images are those of the students; not the tutor. Both solutions demonstrate improvements in the understanding of mobility and greater attention given to presentation. A histogram of the results for the class after marking the second submissions is shown in Fig. 3. This paper has explained why a new method of assignment based on a nonconventional marking key was integrated into the mechanism and multibody dynamics course at UWA"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003120_icate.2012.6403418-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003120_icate.2012.6403418-Figure2-1.png",
        "caption": "Fig. 2. Comparison between the flux density distribution lines for the motor of type M3AA 71 B2/4, 1500 rpm at t=0,5 s for winding : a) Double layer b) Simple layer",
        "texts": [
            " MODELING THE NOISE OF ELECTROMAGNETIC NATURE FOR THREE PHASE INDUCTION MOTOR Changing the speed for this motor is done by changing the number of pole pairs in ratio 2/4 and by changing the connections star / double star. At first it was made a comparison between the distribution of electromagnetic field lines for double layer and single layer winding for the studied motor. Distribution of flux lines in the motor is shown at t = 0.5 s. In Fig. 1 is observed that the field lines at startup are not evenly distributed in the case of double layer winding motor compared with the single layer winding. In Fig. 2 is analyzed the flux density distribution for each type of winding, at t = 0.5 s when the motor function at nominal parameters. Analyzing distributions and legend is observed that the flux density distribution in the case of double layer winding motor has higher values on all engine surfaces, compared with single layer motor windings. 978-1-4673-1810-5/12/$31.00 \u00a92012 IEEE The normal component of airgap magnetic flux density is shown in comparison in Fig. 3 from which results the value of flux density equivalent in the two cases, respectively 1,2 T"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003500_iccas.2013.6704046-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003500_iccas.2013.6704046-Figure4-1.png",
        "caption": "Fig. 4 Gliding locomotion of the developed robot",
        "texts": [
            " One RC servo mo tor, GWS MICROI2BBMG, is used for the actuator of the developed robot, and managed by the wireless communi cation module. Then, we can manually control this servo motor by using the wireless transmitter. And the gliding locomotion robot has a voltage converter and a 7.4V lipo battery. In this paper, we have considered the propulsion method called gliding locomotion, and developed the robot. Then, we show the robot moving principle in this section. We show the principle of gliding locomotion for the proposed robot as Fig. 4. We assume that the model con sisted of two frame-fixed wheel, one passive wheel, one body frame and one arm frame. The arm frame is con nected to rotate with respect to the body frame. The pas sive wheel is attached at the arm frame and kept by the rotary spring to point same direction. Two frame-fixed wheels are attached at the body frame. The proposed robot can move forward to utilize reaction forces and the driving force by a servo motor. Firstly, we consider when the arm frame moves in the same direction as the robot's traveling direction, as shown in Fig. 4 (a). In this time, no force acts on the proposed robot's actuator, except for a friction force. Secondly, we consider the time the arm frame moves in the different as the robot's traveling direction, as shown in Fig. 4 (b). At this moment, the passive wheel keeps pointing same di rection by friction force, and then the force by the rotary spring acts on the wheel. Thus, the passive wheel turns in original direction by this force. As a result, the proposed robot can move forward. We consider about the time that the driving torque by a servo motor acts on the passive joint, as shown in Fig. 5 . nand t represent the normal line direction and the tangent line direction for the wheel rolling direction, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003500_iccas.2013.6704046-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003500_iccas.2013.6704046-Figure2-1.png",
        "caption": "Fig. 2 Mecanical configuration of the developed robot",
        "texts": [
            " In this paper, we have considered mechanism of the gliding locomotion which uses only one servo motor. Moreover, we have developed a robot with this mecha nism which can move forward and turn. Then, we have shown experimental results that illustrate the effective ness of the proposed mechanism. This section contains a brief outline of the mechani cal and electrical configuration of the developed gliding locomotion robot. The gliding locomotion robot has been developed as shown in Fig. 1. The developed robot consists of two frames, a body frame and an arm frame, and three wheels shown in Fig. 2. Two wheels are attached to the body frame and fixed in the moving direction of the robot. One 978-89-93215-05-295560/13/$15 \u00a9ICROS 920 wheel is attached to the arm frame and point in the di rection along the frame. Each frame is connected by the linkage mechanism. One actuator is attached onto body frame. The developed gliding locomotion robot has 300 [mm] length, 150 [mm] width and 68 [mm] height and it weighs 330 [g]. Electrical system diagram for the developed gliding lo comotion robot is described as Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003444_coase.2012.6386360-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003444_coase.2012.6386360-Figure1-1.png",
        "caption": "Fig. 1. Coordinates of belt object",
        "texts": [
            " As the first step of such procedure, we propose a method to identify the flexural rigidity and the initial curvature of a deformable belt object from its static images. Let us assume that a belt object is stuck out horizontally from the edge of a flat table. Then, the object bows in a vertical plane due to the gravity. As the potential energy of the object consists of the flexural and gravitational potentials, the deformed shape of the object is dependent on its flexural rigidity and line density. We introduce the global coordinate system O-xy as shown in Fig.1. Let s be the distance from the origin of O-xy along the object and \u03b8(s) be the angle from the x-axis at point P(s) on the object. Then, the coordinates of point P(s) is described by [ x(s) y(s) ] = \u222b s 0 [ cos \u03b8 sin \u03b8 ] ds. (1) From Fig.1, the following equation must be satisfied as a geometric constraint: \u03b8(0) = 0. (2) In this paper, we assume that a belt object has the linear elasticity. Furthermore, the flexural rigidity, the linear density, and the initial curvature are assumed to be constant at any point on the object. Next, we formulate the potential energy of the object. When a belt object doesn\u2019t have the initial curvature, its potential energy U is described as follows: U = Rf 2 \u222b L 0 ( d\u03b8 ds )2 ds + \u03c1g \u222b L 0 y ds, (3) 978-1-4673-0430-6/12/$31"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003133_0954406211417228-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003133_0954406211417228-Figure11-1.png",
        "caption": "Fig. 11 The crank and slider mechanism",
        "texts": [
            " 8) 10 9, and this range of values was used in the mathematical model of the engine. The piston skirt has been divided into 32 identical pads, each occupying an angle of 45 either side of the centre, as shown in Fig. 2. This number of pads was chosen because no noticeable difference was detected in the results when the number of pads was increased from 16 to 32. The force acting on each of the pads could arise from both elastic and hydrodynamic forces and is determined as follows. Referring to Fig. 11(a) in Appendix 2, the gap between the piston skirt and the cylinder wall at pad YPn is given by EPgap \u00bc GAP xp \u00fe XCYL \u00fe GX \u00fe \u00f0Pa YPn\u00de \u00f05\u00de If EPgap< 0, then EPn \u00bc EPgapKp \u00f06\u00de and HPn \u00bc 0 \u00f07\u00de where EPn is the elastic contact force, HPn the squeeze film force, and Kp the contact stiffness between the piston and cylinder wall. Proc. IMechE Vol. 226 Part C: J. Mechanical Engineering Science at VIRGINIA COMMONWEALTH UNIV on March 13, 2015pic.sagepub.comDownloaded from If EPgap> 0, then EPn \u00bc 0 \u00f08\u00de and HPn \u00bc D1 \u00f0 _xp \u00f0Pa YPn\u00de _ EP 3 gap \u00f09\u00de where D1 is a constant given by D1 \u00bc Sk \u00f010\u00de S is a correction factor, suggested by Cameron [6], to compensate for the side leakage that will result from the pads not being classed as wide",
            " For the same reason, it was decided that the shear deflection could also be ignored. Figure 5 shows the secondary movement of the centre of mass of the piston throughout one complete engine cycle at a speed of 4000 r/min, calculated for the models for single- and four-cylinder engines, in a direction normal to the travel of the piston, there being only elastic contact present (i.e. the original model). Positive translations correspond to displacements of the piston towards the thrust side of the cylinder, which in the model detailed in Fig. 11 in Appendix 2 is the right-hand side, in the positive x direction. The exhaust stroke of the engine is between crank angles 0 and 180 , the induction stroke is between 180 and 360 , compression takes place between 360 and 540 , and the expansion stroke is between 540 and 720 . It can clearly be seen that during the exhaust, induction, and compression strokes, the piston moves from one side of the bore to the other, whereas in the expansion stroke the gas load keeps the piston firmly pressed against the thrust side of the cylinder",
            " APPENDIX 1 Notation Bload torsional load applied to the brake Dn damping coefficients Fload torsional load applied to front flywheel (usually set to zero) g acceleration due to gravity Ib moment of inertia of brake Ic moment of inertia of connecting rod about its centre of mass Ics moment of inertia of crank assembly about its centre of mass Ifl1 moment of inertia of front flywheel Ifl2 moment of inertia of rear flywheel Ip moment of inertia of piston about its centre of mass Kn torsional stiffness of part of crankshaft Kp contact stiffness between piston and cylinder wall Mc mass of connecting rod Mcs mass of crank assembly Mp mass of piston Pload instantaneous load applied to piston crown due to gas pressure Tn torque in part of crankshaft 1 angular position of front flywheel 2 angular position of rear flywheel 3 angular position of brake b coefficient of friction between connecting rod and crank g coefficient of friction between piston and connecting rod m coefficient of friction between crank and crankcase p coefficient of friction between piston skirt and cylinder wall APPENDIX 2 Equations presented here are for the single-cylinder model. Key variables are defined in Fig. 11. The analysis of the four-cylinder model is similar, in that it uses these equations to calculate the motion of each of the cylinder assemblies. The following equations calculate the compressive force and friction torque between the piston and connecting rod (i.e. the gudgeon pin forces) FGx \u00bc \u00f0xc\u00feCRa sin xp\u00feGX PX GY \u00deKpin \u00fe _xc CRa _ cos _xp\u00fe\u00f0GX PX \u00de _ GY _ D2 FGy \u00bc \u00f0yc\u00feCRa cos yp\u00feGY \u00fe \u00f0GX PX \u00de Kpin \u00fe \u00f0 _yc CRa _ sin _ypGY _ \u00fe \u00f0GX PX \u00de _ D2 TP \u00bc gRpin ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi FG 2 x \u00fe FG 2 y q sign\u00f0 _ _ \u00de The following equations calculate the compressive force and friction torque between the connecting rod and the crankshaft (i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002951_amr.323.28-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002951_amr.323.28-Figure1-1.png",
        "caption": "Fig. 1 Structure of the FS Fig. 2 Profile of the WG Fig. 3 Model of the HD",
        "texts": [
            " In consideration of the contact, dynamic load impact, meshing interference between the FS and the circular spline (CS) tooth, and the stress, strain of the FS vary along with motion of the harmonic drive under the nominal load, using Abaqus/Explicit in Abaqus the dynamic simulation is conducted. Finally, the position where FS is prone to arise fatigue fracture is to be predicted, meanwhile, the circumferential meshing rule of the FS tooth is studied. 1.1 Finite Element geometric Model. A HD consists of three components that are a CS with internal teeth, a wave generator (WG) and a FS with external teeth. The FS is key element in a HD, because of its periodic elastic deformation caused by the WG, which its structure is shown in Fig. 1. The profile curve of the WG is: where: R0 is the radius of FS neutral circle; \u03c90 is a constant coefficient of the WG in Fig. 2; \u03c6 is the angle measured from the major axis of the WG. In order to mesh the HDs components reasonably, the following simplifications are taken on the model: (1) The connecting flange between the bottom of FS cup and output axis is removed. Because it is far from the open end of the cup and gear ring, the deformation of the FS is not sensitive with it. The nodes on the ring surface exposed when the connecting flange is removed are coupled to the point on the FS axis, which is convenient for defining boundary conditions in ABAQUS"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002952_irase.4.2013.1.10-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002952_irase.4.2013.1.10-Figure2-1.png",
        "caption": "Fig. 2. Determination the programming position of the tool peak Fig. 3. The axial a of a special worm",
        "texts": [],
        "surrounding_texts": [
            "For manufacturing the special helical channels with the ultraprecision machine tool, it is necessary to solve some specifi c aspects of material removing. For one special worm, having the axial profi le presented in Fig. 3, it necessary to realize a distribution of pressures and chipping energy."
        ]
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.9-1.png",
        "caption": "Fig. 3.9",
        "texts": [
            "8) are also valid in the case of a plane motion. Assuming that the motion takes place in the x, y-plane (cf. Fig. 3.5a), we can write \u03c9 = \u03c9 ez, \u03c9\u0307 = \u03c9\u0307 ez, rAP = r er . Inserting into (3.8), we obtain Equations (3.7a,b) for the position, velocity and acceleration of P : rP = rA + r er, vP = vA + \u03c9 ez \u00d7 r er = vA + r\u03c9 e\u03d5, aP = aA + \u03c9\u0307 ez \u00d7 r er + \u03c9 ez \u00d7 (\u03c9 ez \u00d7 r er) = aA + r\u03c9\u0307 e\u03d5 + \u03c9 ez \u00d7 r\u03c9 e\u03d5 = aA + r\u03c9\u0307 e\u03d5 \u2212 r\u03c92er . 3.1 Kinematics 139 E3.1Example 3.1 A slider crank mechanism consists of a crankshaft 0A and a connecting rod AK (Fig. 3.9a). The crankshaft rotates with a constant angular velocity \u03c90. Determine the angular velocity and the angular acceleration of the connecting rod as well as the velocity and the acceleration of the piston K in an arbitrary position. Solution We choose an x, y-coordinate system and the angles \u03b1 and \u03d5 as shown in Fig. 3.9b. The piston K can move only in the horizontal direction. Therefore, its vertical displacement is zero: yK = r sin\u03b1\u2212 l sin \u03d5 = 0 . (a) This yields the relation sin \u03d5 = r l sin \u03b1 \u2192 cos\u03d5 = \u221a 1\u2212 r2 l2 sin2 \u03b1 (b) between the angles \u03b1 and \u03d5. If we differentiate (a), we obtain the angular velocity \u03d5\u0307 and the angular acceleration \u03d5\u0308 of the connecting rod (\u03b1\u0307 = \u03c90 = const): y\u0307K = r\u03c90 cos\u03b1\u2212 l\u03d5\u0307 cos\u03d5 = 0 \u2192 \u03d5\u0307 = \u03c90 r l cos\u03b1 cos\u03d5 , y\u0308K = \u2212 r\u03c92 0 sin\u03b1 + l\u03d5\u03072 sin \u03d5\u2212 l\u03d5\u0308 cos\u03d5 = 0 \u2192 \u03d5\u0308 = \u2212\u03c92 0 r l sin\u03b1 cos\u03d5 + \u03d5\u03072 sin \u03d5 cos\u03d5 = \u03c92 0 r l [ \u2212 sin \u03b1 cos\u03d5 + r l cos2 \u03b1 sin \u03d5 cos3 \u03d5 ] "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002826_s1644-9665(12)60146-0-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002826_s1644-9665(12)60146-0-Figure2-1.png",
        "caption": "Fig. 2. The constructional solution for the jointed column structure",
        "texts": [
            " The rods are rigidly connected at both ends assuring evenness of movement and the deflection angle at the simply-supported points. Quantity c0, c2 are stiffnesses of the rotational springs at the fixing points of column. In the case of a supporting column at x = 0 fixing stiffness c0 . Euler loaded columns have been studied with two optional end fixing solutions. These are: jointed (both column ends were fixed to rotating loops) and supported (one column end fixed to a rotating loop and the other strengthen). The constructional solution for the jointed Euler loaded column is shown in Figure 2. The active load charging head is the prism 1(1). The head taking the active or passive column 4 load is the set of shaft 2 and accordingly: needle rolling bearing 3(1) \u2013 column A \u2013 or stiff cylindrical element 3(2) \u2013 column B. The head taking the load \u2013 prism 1(2) \u2013 is fixed to base 5. The constructional solution for the supported Euler loaded column is shown in Figure 3. The difference between this solution and the column in Figure 2 is the direct fixing of one end of the column 4 onto base 5. The columns are labelled according to the cylindrical element in use: \u2013 column C \u2013 rolling ball bearing 5(1), \u2013 column D \u2013 stiff cylindrical element 5(2) with circular contour of the working area. A loss of stability occurs in the plane with lower bending stiffness, after exceeding a certain value of the axial force P called critical force Pc. The formulation of the vibration topic is based on the Hamilton rule using the Bernoullie-Euler theory"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure1-1.png",
        "caption": "Fig. 1 Medical air tubing connection system.",
        "texts": [
            " The example comes from a patent published in the year 2011 in the name of \u201c Medical Air Tubing Connection System\u201d [4].The invention is directed to a coupling assembly for use in connecting a supply conduit to a supply source\uff0c 978-1-4673-2237-9/12/$31.00 \u00a92012 IEEE 772 particularly directed to an anti-mismatch coupling assembly and medical air tube connection system for coupling an air supply tube to an air supply manifold to prevent incorrect connection of the supply tube to the required supply source. As shown in Fig.1, inlet end 1 of adapter 5 defines a recessed portion 7 and threaded inner walls 8. Recessed portion 7 has a dimension to be coupled to outlet by the corresponding threads 3 and 8. Inlet end 8 of adapter 5 includes a threaded bore 9 for receiving a threaded screw 10. Threaded bore 9 extends radially through the side of adapter 5 into recessed. This connection\u2019s core concept is that the fluid path 11 goes to the hat 6 directly, while the screw 3 is on the surface of main body 4, which can make sure that no fluid or gas leak"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.22-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.22-1.png",
        "caption": "Figure 3.22 The stability axes, (ie, je, ke), local surface normal, n, and relative flow direction, \u2212iv",
        "texts": [
            " Flow of atmospheric gases relative to the body\u2019s external surface gives rise to aerodynamic force, fv(t), and torque, \u03c4(t), the fundamental sources of both of which are the surface distributions of static pressure, p(x, y, z, t), acting normal to the surface, and shear stress, f (x, y, z, t), along the relative flow direction:4 fv = \u222b A {\u2212pn + f (n \u00d7 [\u2212iv \u00d7 n])} dA, (3.135) \u03c4 = \u222b A d \u00d7 {\u2212pn + f (n \u00d7 [\u2212iv \u00d7 n])} dA, (3.136) 4 Flow far upstream of the body is opposite to the relative velocity vector, v, and is termed freestream. where d(x, y, z) locates an elemental area, dA, with respect to the center of mass (Figure 3.22), n(x, y, z, t) denotes the local surface normal (which is changing in time due to the angular velocity, (P, Q, R)), and relative flow direction is given by \u2212iv, where iv = v v . (3.137) The static pressure and shear stress distributions in a continuous (continuum) flow are themselves functions of the flow velocity far upstream, (v, \u03b1, \u03b2) (also called freestream velocity), the angular velocity of the rigid body relative to the freestream, (P, Q, R), and the thermodynamic properties of the freestream, namely density, \u03c1, static temperature, T , dynamic viscosity, \u03bc\u0304, and specific-heat ratio, \u03b3 ",
            " Based upon Knudsen number, the flow regimes are classified as follows: (i) free-molecular flow, Kn \u2265 10; (ii) transition flow, 0.01 \u2264 Kn \u2264 10; (iii) continuum (or continuous) flow, Kn \u2264 0.01. When the nominal trajectory is a steady maneuver, the aerodynamic forces and moments are balanced by other forces, and thus have constant equilibrium values, fe v , \u03c4 e. In such a case, a special body-fixed frame \u2013 called stability axes \u2013 is generally employed, with one axis initially aligned with the nominal flight direction (Figure 3.22). A small perturbation v, \u03c9 about the nominal trajectory, ve, \u03c9e, v = ve + v, \u03c9 = \u03c9e + \u03c9, (3.138) causes small changes in the flow speed, angle of attack, sideslip angle,5 and the body-rates, (P, Q, R), measured with respect to the stability axes, resulting in the linearized loads by Taylor series expansion: fv = fe v + fv, (3.139) \u03c4 = \u03c4e + \u03c4, (3.140) where fv = 9\u2211 i=1 \u221e\u2211 k=1 \u2202kfv \u2202\u03b6k i \u03b6k i , (3.141) \u03c4 = 9\u2211 i=1 \u221e\u2211 k=1 \u2202k\u03c4 \u2202\u03b6k i \u03b6k i , (3.142) and \u03b6i(t) is the ith element of the relative motion vector, \u03b6(t): \u03b6(t) = ( v, v\u0307, \u03b1, \u03b1\u0307, \u03b2, \u03b2\u0307, P, Q, R )T "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003136_detc2011-48218-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003136_detc2011-48218-Figure1-1.png",
        "caption": "FIGURE 1. DEEP GROOVE SINGLE ROW BALL BEARING MULTIBODY MODEL",
        "texts": [
            " A deep groove single row ball bearing ( SKF 6302) is modeled as a multibody system using the commercial software LMS Virtual Lab Motion [16] in order to simulate the dynamic behaviour of a bearing inserted in a test set-up realized in the Mechanical Department of Katholieke Universiteit of Leuven. The model aims to analyze the influence of different operational conditions on the vibration signals. A further objective consists in the extraction of localized defect effects from the vibration signal (main bearing defect frequencies) and in the simulation of the real vibration signature. First of all a short description of the model is necessary. The bearing model is composed of different bodies of well known geometry dimensions as shown in Fig.1: outer and inner rings, seven spheres and a shaft. Initially the different bodies are considered rigid in order to 2 Copyright c\u00a9 2011 by ASME 2 Copyright \u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 03/18/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use obtain detailed information about the dynamic behaviour of the components at low and medium frequencies. This assumption however limits the correctness of the observations at high frequencies"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003849_20130911-3-br-3021.00114-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003849_20130911-3-br-3021.00114-Figure4-1.png",
        "caption": "Fig. 4. Example of application of the obstacle avoidance algorithm",
        "texts": [
            " After analyzing the state of all sensors it\u2019s generated a vector (vectorsens) where it is signalized the latest active odd sensor and the latest active even sensor. This vector only can have two elements different of zero. The permitted values are 1 and -1. It is 1 if the activated sensor is on the side of the link where the increasing angle of rotation is positive, and it is -1 if it is on the opposite side. The algorithm divides the robots structure in 2 planar structures. One of them is for odd links and the other is for even links. Figure 4 shows a situation when link sensor 5 (or 6) was activated and the correction was performed. The \u2206inc variable given by \u0394inc = sin-1 D/(2L) (8) controls the deviation distance (D) between the obstacle and the link in collision. It is assumed that each link has the length L. The basic idea of this algorithm is to analyze the odd joints and the even joints independently, and according to the activated sensors, the joints vector is update according to ]116[]1616[]116[]116[   sensnext vectorAinc (9) Matrix A in equation (9) is defined as                                                                  1010100000000000 0101010000000000 1010101000000000 0101010100000000 0010101010000000 0001010101000000 0000101010100000 0000010101010000 0000001010101010 0000000101010101 0000000010101010 0000000001010101 0000000000101010 0000000000010101 0000000000001010 0000000000000101 A For example, if the sensor signal of the 7th link is 1 and that of 10th is -1, the final robots configuration is that shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001093_012079-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001093_012079-Figure3-1.png",
        "caption": "Figure 3. Scheme for determining the length of the contact path of a bar with a plane.",
        "texts": [
            "1088/1755-1315/720/1/012079 When using a cylindrical brush, energy is spent both on destroying the contaminant layer and on overcoming additional resistance: aerodynamic, friction in bearings, etc. The work of destruction when separating the contaminant particles from the treated surface due to the friction force of the bar against the contaminant layer can be found from the equation:   xx dPfdXXPfA N x x NC    11 ,)( \u00b7 )( \u00b7 (3) where \u0192 is the coefficient of friction of the bar material on the surface to be treated; \u0420N is the value of the normal pressure of the bar on the surface; \u03b2 is an angle of rotation of the brush core during contact; X is the length of the contact path (Figure 3). The normal pressure force \u0420N is determined from the expression: . )]cos1( \u00b7 [ } \u00b7 ] )cos1( \u00b7 [arccos{ 22 L RdL Vk L Rd kP \u0449VN        (4) International science and technology conference \"Earth science\" IOP Conf. Series: Earth and Environmental Science 720 (2021) 012079 IOP Publishing doi:10.1088/1755-1315/720/1/012079 The drive power expended to destroy the surface of the contaminant by friction can be determined from the expression: , 75 1 AN  (5) where A is bar friction work per second, kgm. The bar operation for one second is defined as: , \u00b7 \u00b7 60 1 \u0449\u041f\u0420C n\u041fAA  (6) where A\u0421 is work of destruction of a bar for one cycle of interaction with a surface; \u041f\u041f\u0420 is the number of brush bars; n\u0449 is the speed of rotation of the cylindrical brush per minute"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002470_amm.397-400.589-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002470_amm.397-400.589-Figure6-1.png",
        "caption": "Fig. 6 Equivalent stress contour of the inner ring",
        "texts": [
            " 2 and Eq. 3. Comparison between FEA results and analytical solutions is shown in Table 1: Table 1 shows that analysis results of these two methods are close and simulation results are in good agreement with analytical solutions .The method that building a FEA model of rolling bearing and then analyzing its dynamic characteristics with ANSYS / LS-DYNA is feasible. Analysis of stress Fig. 4 is slice stress cloud of the bearing at 26ms. Fig. 5 is equivalent stress contour of rolling elements at 26ms. Fig. 6 is equivalent stress contour of the inner ring at 26ms. Fig. 7 is equivalent stress contour of the outer ring at 26ms. Fig. 4 Slice stress cloud of the bearing Fig. 5 Equivalent stress contour of rolling elements Fig. 4 shows equivalent stresses of the bearing are concentrated in the contact areas between the rolling elements and the inner/outer ring. The maximum stresses appear in a certain depth below the contact surfaces and gradually attenuate outwards. Hertz contact theory accounts that: there are compressive stresses in the elliptical contact areas of ball bearings; compressive stresses change in axial and normal directions on the surface; the maximum stresses appear in a certain depth of the contact regions[5]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001756_9781119970422.ch6-Figure6.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001756_9781119970422.ch6-Figure6.5-1.png",
        "caption": "Figure 6.5 Communication with the sink can be guaranteed whenever a vehicle stops to collect data from nodes. (a) Synchronized movement after the 4th visited point. (b) Synchronized movement at the end of the mission. (c) Velocities.",
        "texts": [
            " Adding a simple linear constraint may reduce or extend the computational complexity by magnitudes, without restricting any quality of the solution In the proposed example (Figure 6.4), we minimized the sum of velocities by solving min ui,k ,vi,k { nv\u2211 i=1 nt\u2211 k=1 k ( (vxik) + ( vyik )) } subject to the constraints resulting from (6.1)\u2013(6.5). The number of visited breakpoints per vehicle is at most nt \u2212 1, such that the user is formulate a problem similar to the capacitated CVRP by fixing nt \u2264 nb. Furthermore, in a subsequent computation after solving the MILP, the time intervals [tk, tk+1] can be shortened or stretched according to desired maximal or minimal velocities (Figure 6.5(c)). Due to the NP-hard characteristics of the underlying TSP, the time to compute the guaranteed global optimum explodes with the increasing number of breakpoints and timesteps. For a problem with three mobile entities, the solution of the corresponding MILP3 (342 variables, 1656 constraints) for a setting with 10 points on a grid of seven timesteps took approximately 5.5 s. For a problem with three mobile entities, the computation for a setting with 21 points on a grid of 10 timesteps (816 variables, 8244 constraints) took approximately 16 500 s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002919_acc.2012.6315609-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002919_acc.2012.6315609-Figure2-1.png",
        "caption": "Fig. 2. Aeroelastic model[1]",
        "texts": [
            " These solutions may yield up to three different PWA controllers for the system pwainc. Such stabilizing PWA controllers are guaranteed to stabilize the nonlinear system nlsys contained in the convex hull of pwainc, as shown in [16], [17]. We study airfoil dynamics with plunging and pitch degrees of freedom [1] to show how a nonlinear system can be modeled, analyzed and stabilized through a PWA controller using PWATOOLS. This example has been solved on a machine running Windows XP, Pentium 4 CPU and a memory of 1GB. The model dynamics for the airfoil in Figure 2 is described TABLE III AIRFOIL SYSTEM PARAMETERS M = ( 12.3870 0.4180 0.4180 0.0650 ) C0 = ( 27.43 0 0 0.036 ) C\u00b5 = (31.17 3.99 0.21 \u22120.027 ) K0 = ( 2844.4 0 0 0 ) K\u00b5 = (0 935.1 0 \u22126.3 ) B = (\u2212499.796 \u2212514.680 \u221212.759 \u221214.768 ) \u00b5 = 176.609 in (8) as x\u0307 = ( 0 I \u2212M\u22121(Ko+Ku) \u2212M\u22121(Co+Cu) ) x+ ( 0 \u2212M\u22121 ( 0 x2K\u03b1 (x2) )) + ( 0 \u00b5M\u22121 ) Bu (8) where x= ( x1, x2, x3, x4 )T , x1 = h, x2 =\u03b1 , x3 = h\u0307, x4 = \u03b1\u0307 . The input command is the deflection of the two flaps given by \u03b2 , and the nonlinear function K\u03b1(\u03b1) is defined as follows K\u03b1(\u03b1) = 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003431_2013-36-0272-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003431_2013-36-0272-Figure1-1.png",
        "caption": "Figure 1 \u2013 A rigid body disk",
        "texts": [
            " For a complete revolution of the tire belt, the amount of mass which pass by the point D* must be equal to the tire belt\u2019s Page 4 of 7 total mass M. Using equation (7), we obtain the time required for this complete revolution: (9) As the wheel rim and the tire belt must complete one revolution at the same time, we have from equations (8) and (9): (10) As the total tire belt\u2019s mass M is equal to 0.2 R, we obtain: (11) If we change to the absolute referential shown in Figure 5, we have: (12) where V is the velocity of the non-sliding rigid disk center, as shown in Figure 1 and in equation . Constant acceleration ride case In case of vehicle\u2019s ride in constant acceleration, i.e., wheel rim under constant torque T, the wheel rim\u2019s angular acceleration is constant, as well the relative acceleration and the tire belt\u2019s deformation at point D*. Note that: - as the torque T is constant, the stress (and consequently the strain or deformation ) in the tire belt, when it starts to contact the ground, is always the same (constant); - as there is no sliding between the tire belt and the ground, the deformation will be the same along all the contact extension",
            " For a complete revolution of the tire belt, the amount of mass which pass by the point D* must be equal to the tire belt\u2019s total mass M. Using equation (7), we obtain the time required for this complete revolution: (14) As the wheel rim and the tire belt must do a complete revolution at the same time, we have from equations (13) and (14): (15) As the total tire belt\u2019s mass M is equal to 0.2 R, we obtain: (16) If we change to the absolute referential shown in Figure 5, we have: (17) where is the acceleration of the non-sliding rigid disk center, as shown in Figure 1 and in equation . If we multiply both sides of equation (17) by the time t, we obtain: (18) Note that this relationship is the same of the case of constant velocity, in equation (12). In cases when the acceleration is not constant, but it changes smoothly as in usual urban or highway trips, the equation (18) (or (12)) still may be used without significant loss of accuracy. Next, we will examine the three non-sliding \u201cslip\u201d cases: longitudinal, lateral and spin. In the case of a rigid tire, which preserve its circular shape even under vertical loading, as shown in Figure 1, or in the case of a flexible but inextensible tire belt, the \u201cinextensible relationship\u201d between the velocity Vi of the wheels\u2019 center O and its rotational velocity , as seen before, is given by: Page 5 of 7 (19) where R is the external radius of the set (wheel + tire). When the tire is vertically loaded, the geometry of the tire belt region, in contact with the ground, changes from a cylindrical to a plane shape. Even so, if the tire belt is inextensible, the equation (19) remains valid. But, as the tire belt is in traction due to the internal tire pressure, the contact region suffers a \u201cstress relief\u201d due to that geometric change, while the remaining of the tire belt - not in contact with the ground - preserves its cylindrical shape"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003876_aqtr.2012.6237754-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003876_aqtr.2012.6237754-Figure3-1.png",
        "caption": "Figure 3 \u2013 Active Instrument \u2013 Actuators position",
        "texts": [
            " PARASURG 5M has 5 actuators, three of them are installed at the base platform of the robot and the other two integrated in the orientation module structure (Fig.1). The structure of PARASURG 5M is build in such a way that PARASIM can be easily interconnected [8]. PARASIM robotic instrument model was developed using CAD, integrating a spherical guiding mechanism (Fig.2) The system is actuated by a series of DC motors, and motion transmission elements (belts, gears) in order to transmit the 3+1 degrees of freedom: q6, q7 and q8 (rotations) as well as one translation \u2013 q9 (acting as a griper). Fig.3 shows the position of the actuators which control the end effector. The elements of the instrument are the driving wire, speed reducer, the toothed belt, belt wheel and the rod. The actuation of the robot is achieved using nine electrical motors, produces by Maxon company [17]. The robot uses actuators to perform the following tasks: - Positioning of the arm by actuators M1, M2, M3 - MCD EPOS 60 W type; - Guiding the robot tip by actuators M4, M5 \u2013 EC 32 flat, brushless, 6 W, with Hall sensors; - Instrument tip motion by actuators: M6\u2013M9 \u2013 EC 6,1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.113-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.113-1.png",
        "caption": "Fig. 2.113 4PaRPaRR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 26, limb topology Pa\\R\\Pa\\kR\\R",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.20-1.png",
        "caption": "Figure 3.20 The aircraft body axes, (i, j, k), relative to the NED local horizon frame, (I, J, K)",
        "texts": [
            " Thus, the orientation of an aircraft\u2019s body-fixed frame, (i, j, k) \u2013 where (i, k) is the plane of symmetry \u2013 relative to an inertial, north, east, down (NED) local horizon frame, (I, J, K), where I = i\u03b4, J = i , (3.111) K = \u2212ir, Table 3.5 % Program for specifying governing ODEs of great circle navigation % expressed as state equations (to be called by \u2019ode45.m\u2019) % (c) 2010 Ashish Tewari function dydx=airnav(x,y) global vprime; global vw; global Aw; global r; global Omega; v=sqrt(vprime\u02c62-vw\u02c62*sin(Aw-y(3))\u02c62)+vw*cos(Aw-y(3)); dydx=[v*cos(y(3))/r v*sin(y(3))/(r*cos(y(1))) v*sin(y(3))*tan(y(1))/r]; is expressed through the Euler angles, ( )3, ( )2, (\u03c3)1 (Figure 3.20), as\u23a7\u23a8 \u23a9 i j k \u23ab\u23ac \u23ad = C \u23a7\u23a8 \u23a9 I J K \u23ab\u23ac \u23ad . (3.112) Here, C is the rotation matrix C = \u239b \u239d cos cos cos sin \u2212 sin sin \u03c3 sin cos \u2212 cos \u03c3 sin sin \u03c3 sin sin + cos \u03c3 cos sin \u03c3 cos cos \u03c3 sin cos + sin \u03c3 sin cos \u03c3 sin sin \u2212 sin \u03c3 cos cos \u03c3 cos \u239e \u23a0 , where \u03c3 is the roll (or bank) angle, the pitch angle, and the yaw angle. The angular velocity of the rigid aircraft about its center of mass is given by \u03c9 = P i + Qj + Rk, (3.113) where P is the roll rate, Q the pitch rate, and R the yaw rate. Using the Euler angle representation ( )3, ( )2, (\u03c3)1 as shown in Figure 3.20, we have the rotational kinematics equation \u03c9(t) = \u03c3\u0307i + \u0307J\u2032 + \u0307K, (3.114) which, considering the elementary rotations involved, becomes \u03c9(t) = \u23a7\u23a8 \u23a9 P Q R \u23ab\u23ac \u23ad = \u23a7\u23a8 \u23a9 \u03c3\u0307 \u2212 \u0307 sin \u0307 cos \u03c3 + \u0307 sin \u03c3 cos \u2212\u0307 sin \u03c3 + \u0307 cos \u03c3 cos \u23ab\u23ac \u23ad (3.115) or \u23a7\u23a8 \u23a9 \u03c3\u0307 \u0307 \u0307 \u23ab\u23ac \u23ad = 1 cos \u239b \u239d cos sin \u03c3 sin cos \u03c3 sin 0 cos \u03c3 cos \u2212 sin \u03c3 cos 0 sin \u03c3 cos \u03c3 \u239e \u23a0 \u23a7\u23a8 \u23a9 P Q R \u23ab\u23ac \u23ad . (3.116) It is evident from equation (3.116) that the Euler angle representation given by the sequence ( )3, ( )2, (\u03c3)1 is singular for = \u00b1\u03c0/2. However, since an aircraft normally does not approach the vertical attitude, this particular Euler angle sequence is commonly applied to aircraft stability and control applications",
            " The difference between the usage of the two sets of equations is that the mass is assumed constant during the relatively small time scale of attitude dynamics, while the bank angle and the angle of attack \u2013 variables of attitude dynamics \u2013 are taken to be either constants or external inputs in the navigational plant. The external torque vector generated by aerodynamic and propulsive means is expressed for an aircraft as \u03c4 = Li + Mj + Nk, (3.122) where L is the rolling moment, M the pitching moment, and N the yawing moment (Figure 3.20). All aircraft have at least one plane of symmetry, which we shall denote here by the plane, oxz, bound by the body-fixed axes, (i, k), as shown in Figure 3.20. By virtue of the plane of symmetry, we haveJxy = 0 = Jyz and the rotational kinetics equation can be expressed as \u03c4 = \u23a7\u23a8 \u23a9 L M N \u23ab\u23ac \u23ad = J\u03c9\u0307 + S(\u03c9)J\u03c9, (3.123) where S(\u03c9) = \u239b \u239d 0 \u2212R Q R 0 \u2212P \u2212Q P 0 \u239e \u23a0 (3.124) and J = \u239b \u239d Jxx 0 Jxz 0 Jyy 0 Jxz 0 Jzz \u239e \u23a0 . (3.125) Substituting equations (3.124) and (3.125) into equation (3.123) and collecting terms, we have the following scalar equations of rotational kinetics: L = JxxP\u0307 + Jxz(R\u0307 + PQ) + (Jzz \u2212 Jyy ) QR, (3.126) M = JyyQ\u0307 + Jxz(R 2 \u2212 P2) + (Jxx \u2212 Jzz) PR, (3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002529_mechatron.2011.5961097-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002529_mechatron.2011.5961097-Figure4-1.png",
        "caption": "Fig. 4. The scheme, illustrating the way of decreasing the angle \u03b1",
        "texts": [
            " Moreover, the advantage of this scheme will be the application of the first phalanx for a rather wide range of diameters of the object, which is impossible in case of other enumerated schemes. Similar to the general case, the angle \u03b1_\u04401 will depend on many parameters. The more acute the angle will be, the more effectively the gripped object will be controlled. It will be ensured by a smaller rolling friction coefficient, a smaller sliding friction coefficient, a lower radius of the workpiece, a greater length of the first phalanx and a greater torque at the drive of the phalanx. The angle \u03b1_\u04401 can also be reduced structurally, for example, as it is shown in Fig. 4. The quality of mating and the efficiency of operations of transportation and positioning will be influenced by many factors: the sliding friction force in the contact point; angles of mating \u03b1 and \u03b2 (the greater these angles are, the more efficient the gripping process will be and under lower stresses); layout features of the gripper (number and lengths of phalanxes and fingers); properties of the material (of both gripping elements of the gripper and of the object itself); and torques in drives of finger phalanxes (the higher the torques will be, the more reliable the gripping process will be, however, at the same time the stress in the contact will be increasing)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001817_9781118443293.ch7-Figure7.24-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001817_9781118443293.ch7-Figure7.24-1.png",
        "caption": "Figure 7.24 (a) Impedance plane plot for a microelectrode, (b) modified Randles equivalent circuit for a microelectrode to achieve improved mobilities.",
        "texts": [
            "23 shows the expected changes in EIS spectra for an electrode in a bare state (a), then with a sensing layer (b) and finally after antibody attachment (c), as described in [1]. Microelectrodes do not show the same response as macroelectrodes, particularly at low frequencies where hemispherical diffusion means there are no reaction limiting diffusion issues. This requires a modified Randles circuit with a resistor (Rnl) in parallel with the Warburg impedance. This means that the impedance will tend towards a constant resistance at very low frequencies, which represents the stable current seen at dc with a microelectrode. Figure 7.24 shows the modified Randles circuit along with a model impedance plane plot for this type of microelectrode experiment. The Metal Oxide Semiconductor Field Effect Transistor (MOSFET) is one of the most ubiquitous electronic components in the world with most modern computing devices containing several billion. Although most electrical engineers will have at least some familiarity with \u2212I m (Z ) \u2212I m (Z ) the theory of operation of the MOSFET it is worth reviewing the important characteristics of the device before looking at how this structure can be used to implement biological and chemical sensors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003226_icuas.2013.6564716-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003226_icuas.2013.6564716-Figure1-1.png",
        "caption": "Fig. 1. A configuration of the CNU-ducted fan UAV",
        "texts": [
            " Section IV reports the numerical simulation environment and simulation results for the full flight condition from hover to cruise modes. Section V concludes the paper with future work. A configuration of the CNU-ducted fan UAV is introduced in this section. Also, precise linearized modeling data of the CNU-ducted fan UAV at each flight condition is established using mathematical approach which is divided into three modes with respect to airspeed: hover (0m/s), transition (5 and 10m/s) and cruise mode (15m/s). A configuration of the CNU-ducted fan UAV is conventional ring-wing type as shown in Fig. 1. It has four control surfaces that are located at the end of the duct. Also, it contains fixed stators for anti-torque and additional lift. A fuselage is in center of the vehicle, and avionics is mounted in the duct or the fuselage. In addition, payload bay is placed on top of the fuselage. For various missions, operating equipments such as camera, spot light and communication relay can be located at this bay. A coordinate system shown in Fig. 2 has dynamic features similar to a helicopter: thrust vector, anti-torque, gyroscopic coupling, and velocity induced by a main rotor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003644_amr.605-607.1176-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003644_amr.605-607.1176-Figure4-1.png",
        "caption": "Fig. 4 The Model of Eccentric Shaft with Constrain",
        "texts": [
            " In practical, the bearing is not rigid support,but the elastic. In order to decrease the error,the bearings are regarded as eight springs which are built by combination14, and the stiffness of springs are regarded as the radial direction of bearing. In addition,add 16 nodes on the out of bearing and creat the line between the point on the model of bearing and the point created just now, and build the model of spring.The full constraints are exerted on the one side of spring,and axial constraint is exerted on the other side. The result are shown in the Fig.4. In order to study the influence between the natural frequency and elastic constraint,the stiffness of springs random from 3.96x10 7 N/m,3.96x10 8 N/m,3.96x10 9 N/m,3.96x10 10 N/m,3.96x10 11 N/m,3.96x10 12 N/m [4] and the different natural frequency are obtained at the different stiffness.The result are shown in Table.2. Table. 2 The Natural Frequency of Eccentric Shaft at Different Stiffness Stiffness(N/m) Natural frequency (Hz) Critical speed (r/min) 3.96E+07 3.96E+08 3.96E+09 3.96E+10 3.96E+11 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.6-1.png",
        "caption": "Figure 3.6 The relative velocity vector, v\u2032, in the presence of wind velocity, vw, and the modified wind axes, i\u2032v, j\u2032v, kv",
        "texts": [
            " The ground track, on the other hand, is affected even by a steady wind. Let the aiplane\u2019s velocity relative to the rotating planet be v, while a horizontal wind of velocity vw = vw(sin Awi + cos Awi\u03b4) (3.30) 2 The jet stream is a very strong west-to-east \u201criver of air\u201d confined to a narrow altitude range near the tropopause and has a width of a few kilometers, but covers almost the entire globe, with core wind speeds reaching 500 km/hr. is blowing at a steady speed, vw, and velocity azimuth, Aw, as shown in Figure 3.6. The airplane\u2019s velocity relative to the atmosphere, v\u2032, is given by v\u2032 = v \u2212 vw = v sin \u03c6ir + (v cos \u03c6 sin A \u2212 vw sin Aw)i + (v cos \u03c6 cos A \u2212 vw cos Aw)i\u03b4, (3.31) and the effective sideslip angle, \u03b2, due to the wind is \u03b2 = A \u2212 A\u2032 (3.32) (Figure 3.6), where A\u2032 is the airplane\u2019s velocity azimuth relative to the atmosphere. The coordinate transformation between the modified wind axes, (i\u2032v, j\u2032v, kv), and the original (zero-wind) wind axes, (iv, jv, kv), is given according to Figure 3.6 by{ iv jv } = ( cos \u03b2 sin \u03b2 \u2212 sin \u03b2 cos \u03b2 ){ i\u2032v j\u2032v } . (3.33) One can express the airplane\u2019s relative velocity in the modified wind axes as v\u2032 = v\u2032i\u2032v where the airplane\u2019s airspeed is given by v\u2032 = \u221a v2 + v2 w \u2212 2vvw cos \u03c6 cos(A \u2212 Aw), (3.34) according to the velocity triangle depicted in Figure 3.6. The governing kinematic equations (3.19)\u2013(3.21) are modified by the wind in so far as the ground speed, v, changes due to the wind velocity, (vw, Aw), and the velocity azimuth, A, according to equation (3.34). Similarly, the aerodynamic and propulsive forces (Figures 3.3 and 3.5) with the approximation \u03b5 \u03b1 \u2013 resolved in the axes iv, jv, kv in equation (3.14) \u2013 are changed due to the change in airspeed (equation (3.34)) and the wind axes, i\u2032v, j\u2032v, kv. Let the modified aerodynamic and propulsive forces in a coordinated flight be expressed as follows: Fv = ( f \u2032 T cos \u03b1\u2032 \u2212 D\u2032) i\u2032v + L\u2032 sin \u03c3j\u2032v \u2212 (f \u2032 T sin \u03b1\u2032 + L\u2032 cos \u03c3 ) kv = (fT cos \u03b1 \u2212 D + fXw)iv + (L sin \u03c3 + fYw)jv \u2212 (fT sin \u03b1 + L cos \u03c3)kv, (3",
            " In summary, we can write the state equations for the two subsystems as follows, and represent them by the block diagrams shown in Figure 3.7: (a) \u03b4\u0307 = v cos A R0 + h , (3.43) \u0307 = v sin A (R0 + h) cos \u03b4 , (3.44) A\u0307 = L sin \u03c3 + fYw mv + v r sin A tan \u03b4 + 1 2v 2r sin A sin 2\u03b4 + 2 sin \u03b4, (3.45) and either m\u0307 = \u2212 cTfT g0 (3.46) or m\u0307 = \u2212 cPfTv g0 . (3.47) Here, the ground speed, v, is related to the constant airspeed, v\u2032, and the constant wind velocity, (vw, Aw), according to the velocity triangle shown in Figure 3.6 as follows: v\u2032 = \u221a v2 + v2 w \u2212 2vvw cos(Aw \u2212 A) (3.48) or v = \u221a (v\u2032)2 \u2212 v2 w sin2(Aw \u2212 A) + vw cos(Aw \u2212 A). (3.49) (b) h\u0307 = v\u2032\u03c6, (3.50) v\u0307\u2032 = fT \u2212 D m \u2212 g0\u03c6 (3.51) \u03c6\u0307 = fT\u03b1 + L cos \u03c3 \u2212 mg0 mv\u2032 , (3.52) where \u03c3(t), m(t) are the reference solutions for subsystem (a). The design of an automatic navigation system thus consists of determining and following a nominal trajectory for subsystem (a) which is then provided as a reference to be followed by subsystem (b). Consequently, while both terminal and tracking control are necessary for subsystem (a), only a regulator needs to be designed for subsystem (b) for maintaining a constant airspeed and altitude"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002487_s11044-013-9384-5-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002487_s11044-013-9384-5-Figure2-1.png",
        "caption": "Fig. 2 Introduced degrees of freedom q1\u2013q9, forces c1\u2013c9, and the reaction forces (internal forces) at oar locks, feet, and seat",
        "texts": [
            " The arm was allowed to have out-of-plane motion through two rotational dofs at the shoulder (q5, q6), and one rotation at the elbow (q7). The two oars, modelled as one unit, were given two rotational dofs, (q8, q9). Five control forces, i.e. torques were used to drive the model and act in the physical joints: ankle, knee, hip, shoulder (flexion/extension), and elbow. Four coupling forces were used to represent the conditions at the slider (c4), and the three force components at the oar handle (c7, . . . , c9), Fig. 2. As the two oars were modelled as one unit, c7, . . . , c9 represent the added force components at both handles. As baseline physical measures for the model were chosen L1 = L2 = L3 = L5 = 0.5 m, L4 = 0.4 m, Linboard = 0.88 m, Loutboard = 1.96 m, Hoar = 0.215 m, Xoarlock = 0.25 m, Yoarlock = 0.64 m, Zoarlock = 0.36 m, HFS = 0.18 m, HGAW = 0.26 m. The masses of the rower\u2019s segments were Mfeet = 4 kg, M1 = 8 kg, M2 = 16 kg, M3 = 32 kg, M4 = 6 kg, M5 = 6 kg, representing the masses of both right and left limbs\u2019 segments"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.63-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.63-1.png",
        "caption": "Fig. 2.63 4PaPRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\P||R\\Pa",
        "texts": [
            "22o) Idem No. 4 43. 4PRPaPa (Fig. 2.55) P\\R\\Pa||Pa (Fig. 2.22p) Idem No. 5 44. 4PPaRPa (Fig. 2.56) P||Pa\\R\\Pa (Fig. 2.22q) Idem No. 33 45. 4PPaPaR (Fig. 2.57) P||Pa||Pa\\R (Fig. 2.22r) Idem No. 4 46. 4PaPPaR (Fig. 2.58) Pa||P||Pa\\R (Fig. 2.22s) Idem No. 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.2-1.png",
        "caption": "Figure 2.2. The developed diagram of",
        "texts": [
            " At any instant the curve of flux density around the air gap circumference may be of any form and is not necessarily sinusoidal. Themain flux of an ACmachine is defined to be that determined by the fundamental component of the curve of air gap flux density, and the radial line where the fundamental density is a maximum is called the axis of the flux. The main flux is then completely defined by a magnitude and a direction [2]. In order to calculate the flux due to a given system of currents it is first necessary to determine the magnetomotive force (mmf) due to the currents. Figure 2.2 is a developed diagram for a two-pole machine extending between angular position 0 and 2p. The currents in the conductors of a coil are distributed in slots as indicated and form two bands, symmetrically distributed about the points A and B. The currents in two bands flow in opposite directions. Since the current distribution is known, the mmf round any closed path can be found; in particular, the mmf round a path crossing the air gap, such as ACDFGH. Because of the high permeability of the iron, all the mmf round the closed path can be assumed to appear across the air gap at the point F, if A is chosen as a point of zero air gap flux density, which, for a symmetrical distribution of current, must occur at the center of each band",
            " Hence, a curve of mmf distribution round the air gap can be drawn for any value of current flowing in the coil. Thus, although magnetomotive force is fundamentally a line integral round a closed path, a value can be associated with each point along the air gap circumference, giving the space distributed mmf curve of the machine. If the conductors are assumed to be located at points around the air gap circumference, the mmf curve is a stepped curve, but the fundamental component, which, by symmetry, is zero at the points A and B, can be drawn as in Figure 2.2. The radial line at the point of maximum mmf (x\u2013x in Figure 2.2) is called the axis of the mmf, and since this depends only on the conductor distribution, it is also the axis of the coil. The curve shows the instantaneous magnitude of the mmf, which depends on the instantaneous value of the current. The mmf curve determines the flux density curve: If the machine has a uniform air gap, and if saturation is neglected, the flux density is everywhere proportional to the mmf. In a salient pole machine, because the flux density is not proportional with mmf, in order to calculate the flux, it becomes necessary to resolve the mmf wave into component wave along the direct and quadrature axes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure22-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure22-1.png",
        "caption": "Figure 22 FDM candidate fixture part (see online version for colours)",
        "texts": [
            "00 4 Sheet metal screws Commercial item 0.25 1.00 Total 1,101 1 week Another case example involves a need for an aerospace manufacturer to construct a tooling fixture for testing a composite part in elevated temperature regimes. The fixture must be fabricated quickly, as a test schedule has been compressed to meet tight deadlines. The fixture must be capable of withstanding a 93\u00b0C sustained operating temperature and offer accuracy comparable to basic fixture tooling. The fixture geometry is shown in Figure 22. FDM was chosen due to the ability to build parts with high complexity quickly. A higher temperature capable material, PPSF, was chosen to accommodate the temperature requirements. In addition, due to the geometry flexibility, design engineers engraved text work instructions and quality inspection requirements directly into the computer-aided design model to aid test technicians in integrating the part in the test, as shown in Figure 23. To summarise, in order to justify RM from an economic standpoint, often design integration becomes a key element"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002689_kem.488-489.194-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002689_kem.488-489.194-Figure3-1.png",
        "caption": "Fig. 3 Residual stress measurement of railway wheel",
        "texts": [
            " In addition, the replication tests were conducted to compare the left and right side wheels with reference to the direction of running. Fig. 1 shows the inspection points of the friction surface of a thermally damaged wheel. 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: 128.42.202.150, Rice University, Fondren Library, Houston, USA-19/05/15,21:09:48) Fig. 2 ~ Fig. 3 show the results of the replication inspections on the friction surfaces of the power car wheels. In Fig. 2(a), Friction Surface-Section 1 on the outer side of the wheel is the part where the deformation by the friction with brake block is the severest. This section 1 showed development of reticular cracks in the white layer formed on the wheel surface caused by rapid heating close to the melting point and cooling. The Friction Surface-Section 1 seems to have been affected by heat generated from friction with brake block most severely, which was the major cause of the severe thermal discoloration on the side",
            " It was assumed that the main causes of the wear and damage of the wheels were the friction between the wheel and brake block and the wheel and rail for the inner and outer sides of the wheel, respectively. It was also assumed that the partial difference of the deformation layer was due to the uneven friction forces between the wheel and brake block and rail. The boundaries between the deformation layers showed micro-cracks vertical and horizontal to the axis of friction. The residual stresses of the damaged and undamaged wheels of high speed train were measured on the tread within 20 mm from the rim, as shown in Fig. 3. Residual stress can be measured by punching method, cutting and strain gauge method, or ultrasonic method. In this study, an X-ray measurement system was used to measure the residual stress of wheels of power car, as shown in Fig. 3(a). The measurement points were arranged at the intervals of 90\u00b0 in the circumferential direction, and the X-ray irradiation angle was -45\u00b0~45 at 25\u00b0 intervals. Table 1 illustrates the measurement conditions of the residual stress of the wheels. The measurement apparatuses of the residual stress of the wheels were XSTRESS 3000 (Stresstech, Finland). The direction of measurement was in circumferential direction. Regarding the diffraction condition of the X-ray, the shifts of diffraction peak were measured at the diffraction angle of , according to the stress measurement method, and tilted in both \u00b1 directions for 5 times ( = 0\u00b0, 20"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure5-1.png",
        "caption": "Figure 5 FDM tensile coupon (see online version for colours)",
        "texts": [
            " Therefore, multiple tensile bars must be built at the same time and attached together to form a larger structure that is not as susceptible to the vibration inherent in the FDM process. Given that FDM Z-axis tensile properties are generally lower than X\u2013Y tensile properties, design engineers must design to the weakest parameter for direct manufacturing of parts. Therefore, a test specimen must be designed to optimise the FDM system relative to Z strength. To address this need in the most efficient possible manner, a test specimen was designed to capture the structural integrity and robustness of the FDM process. The specimen geometry shown in Figure 5 is a grouping of ASTM D638 Type I tensile specimens oriented in the Z-direction. This grouping was originally designed in Dassault Systemes CATIA package, imported in Magics and then exported to the FDM software, Insight. For this example, a FDM 400MC is used to construct the specimen grouping. However, the tensile specimen grouping may be constructed from any FDM material. The ASTM D638 Type I callout describes a tensile specimen that is illustrated in Figure 6. The geometry highlighted in Figure 5 serves several functions. One function replicates the layered pattern of any part built concurrently in the FDM process. The arched geometry connecting tensile specimens acts as a self-supporting feature to eliminate the need for support material to be constructed for the tensile bar grouping. Without the arching feature, the requirement for support material would double the estimated build time. In addition, by eliminating the need for support material, the specimen requires 45% less material. This connecting arched geometry also acts as a rigid body for the construction of the tensile bar group to prevent a tensile bar from toppling over during building",
            " If the stresses in the laminate axes are denoted by x, y and xy, then these are related to the stresses referred to the material axes by the usual transformation equations: 2 2 1 2 2 2 2 2 12 2 2 x y xy c s cs s c cs cs cs c s (1) where c denotes cos and s denotes sin . Also, the strains in the material axes are related to those in the laminate axes, namely x, y and xy, by what is essentially the strain transformation: 2 2 1 2 2 2 2 2 12 2 2 x y xy c s cs s c cs cs cs c s (2) Consider a similar approach using FDM, whereas the tensile specimen highlighted in Figure 5 is shown from a top view perspective in Figure 8. Raster orientations are additively constructed in the FDM process and specified as toolpaths in the Insight software. Therefore, a natural expansion of thought would be to consider that each ply in composite laminate theory may be thought of as a raster pattern layer in the FDM process and by specifying each layers orientation to 0\u00b0 each ply or layer, the specimen may be defined to serve as a function of loading for multi-directionally loaded parts"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001944_9781118354162.ch13-Figure13.22-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001944_9781118354162.ch13-Figure13.22-1.png",
        "caption": "Figure 13.22 Diffusion geometry at a millimeter-sized electrode (A) and at an ultramicroelectrode (B). Adapted with permission from [10]. Copyright 2002 A. Ba\u0306nica\u0306 and F.G. Ba\u0306nica\u0306.",
        "texts": [
            " SWV curves display a large and potential-dependent background that distorts the SWV peak and damages the accuracy of the peak current. Despite its reputed sensitivity, SWV does not perform well in this case owing to the irreversible character of the electrochemical reaction. RDCP appears as a suitable alternative in such situations. Traditional electroanalytical chemistry makes use of working electrodes with sizes in the millimeter region. The very small area of these electrodes leads to very small currents (in the mA region) which do not alter the bulk concentration of the analyte during the run. As shown schematically in Figure 13.22A, the diffusion layer developed at such an electrode is much thinner than the electrode size. Consequently, even if the diffusion paths converge to the center, the effect of convergence can be neglected as the diffusion paths within the diffusion layer are nearly parallel. A good analogy is the fall of two objects from a height much lower than the Earth radius. Although the fall paths converge to the Earth center, the convergence effect is negligible and the objects\u2019 paths are practically parallel",
            " As a consequence of the diffusion conditions discussed above, the linear diffusion model is a very good approxi- mation to the diffusion process at nonplanar electrodes, such as spherical or cylindrical electrodes. The situation changes fundamentally when the electrode size is in the micrometer region or lower. Such electrodes are termed ultramicroelectrodes or, more simply, microelectrodes [25,26]. In this case, the thickness of the diffusion layer becomes much greater then the electrode size and the convergence of the diffusion paths cannot be neglected (Figure 13.22B). Ultramicroelectrodes are widely used in the design of very small sensors for applications in the analysis of microsamples or in in vivo sensing. The particular features of ultramicroelectrode electrolysis will be first illustrated for the case of a spherical ultramicroelectrode. A theoretical approach to diffusion at a spherical electrode demonstrates that the current produced in response to a potential step (under the limiting-current conditions) is given by the following equation: i \u00bc nFDAc r 1\u00fe r pDt\u00f0 \u00de1=2 \" # \u00f013:95\u00de Here, r is the electrode radius, c is the reactant concentration and the other symbols have their usual meaning"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003302_amr.487.327-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003302_amr.487.327-Figure5-1.png",
        "caption": "Fig. 5 The results of the surface mesh Fig. 6 Virtual mold",
        "texts": [
            " The workload of the meshing method can be greatly reduced through the combined use of Geomesh and Meshcast. The efficiency has been increased and the functionality of the software has been enhanced. To a certain extent, the complexity of the establishment of the finite elements is improved. For the establishment of the precision casting model, its shortcut and automatically are absolutely no less than the other software of differential element, but higher than the accuracy of the differential element [3]. The air intake hood after meshing is shown in Figure 5. The virtual mold needed to be set to predict the filling and solidification results and the main defects which were generated in the process. The block in Figure 6 is the virtual mold. After that, the materials, contact conditions, boundary conditions, gravity, initial conditions and the operating parameters are defined respectively to complete the pre-treatment of the numerical simulation of the filling and the solidification process of the air intake hood [4][5][6]. Then, all the defined parameters were saved and the dynamic simulation of the filling and solidification process of the air intake hood could be operated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002431_amr.569.620-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002431_amr.569.620-Figure7-1.png",
        "caption": "Fig. 7 Maximum stress of the pad by 3 rd analysis method",
        "texts": [
            " In addition, the rotational degree of freedom in the circumferential direction of the pads has been analyzed according to the boundary condition by considering the hydraulic oil. The results of the pad analysis give the maximum deformation of 0.0126mm and the maximum stress value of 78.1MPa. By these analytical methods, the actual value appeared similar to the analysis results. The stress distribution shape expected of the initial design was similarly observed. Thus, the pad modeling and method of applying the analysis to the actual boundary conditions is ideal. Fig. 6 and Fig. 7 show the total deformation and equivalent stress of the pad. Tilting pad bearing case considered with boundary condition A tilting pad bearing can be used more efficiently depending on how the same bearing is installed. Therefore, the analysis method of the bearing pad or bearing according to the boundary condition is important because it determines the efficiency. Therefore, in order to apply boundary condition most similar method of actual installing bearing case, it is divided into two situations 4 pads and 5 pads used to built in three dimensional modeling and finite element analysis model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003638_icredg.2012.6190454-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003638_icredg.2012.6190454-Figure2-1.png",
        "caption": "Fig. 2. A view of MFC structure",
        "texts": [
            " Cathode surface was coated with platinum as catalyst (10% Pt, Volcun XC-72, Premetek) for reaction in cathode chamber. Aeration was carried out using a small air pump, and a handmade sparger was set up in the cathode compartment for uniform distribution of air in catholyte. A flow of nitrogen gas (>99% purity) was bubbled into the anolyte and exhausted from head space to keep the anode chamber anaerobic. Electrodes were 978-1-4673-0665-2/12/$31.00 \u00a92012 IEEE 158 connected together using stainless steel wires, tightly attached to the electrodes, and a 10 K\u03a9 potentiometer to set the resistance. Figure 2 shows a view of MFC structure used in this study. Shewanella sp. was purchased from Persian Type Culture Collection (PTCC 1711). Before inoculation, cells were grown in Tryptone Soya Broth (TSB) growth medium for about 70 hours at 30 C (shaking at 140 rpm, aerobically). The MFC medium consisted of the following components [7]: PIPES buffer, 15.1 g/l; sodium hydroxide, 3 g/l; ammonium chloride, 1.5 g/l; potassium chloride, 0.1 g/l; potassium phosphate, 0.6 g/l; sodium chloride, 5.8 g/l; L-glutamic acid, 1",
            " Voltage and current of the system were recorded on-line and power, external resistance, current density, and power density were calculated based upon. All experiments were carried out at C 130 inside an incubator. III. RESULTS AND DISCUSSION A. Voltage generation Once the MFC was inoculated current generation was observed. However, because of the lack of biofilm on the anode surface the current and voltage data were quite unstable. For the first four days, the MFC was periodically fed with inoculant and fresh medium until the cell voltage adopted values in a certain interval (Figure 2a). On day 5, the MFC reached its maximum voltage and was stable for nine consecutive days (Figure 2b). With changing the pattern of inoculation and feeding, the MFC voltage underwent fluctuations. However, the overall voltage lowered to one-fifth over the next 16 days (Figure 2c). Upon 30th day, the voltage neither dropped nor increased noticeably until the last day of experiment (Figures 2d and 2e). The stability of current at this stage was a good sign of forming biofilm on the anode surface. Mature biofilms tend to a sustainable current as long as the substrate is available. A polarization curve is used to characterize current as a function of voltage. By changing the circuit external resistance, we obtain a new voltage, and hence a new current at that resistance [8]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003384_amm.344.3-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003384_amm.344.3-Figure1-1.png",
        "caption": "Fig. 1 Schematic plan of joint wear calculation of ACSPB under radial load",
        "texts": [
            " It provided a very useful method to calculate and forecast the wear-life accurately, and it would be very beneficial to the reliability research on SPB. In order to get the wear-life model conveniently, we made the following simplifications of the working conditions of ACSPB on the premise that the wear mechanism and failure mode would not change. Firstly, the load acting on the ACSPB was mainly the radial load (pure radial load or the composite load of radial load and thrust load while the radial load was the chief load). The radial load acted on the outer ring of ACSPB (as shown in Fig.1 (a)). Secondly, the ACSPB\u2019s movement was only the oscillation along the axial lead, ignoring all other oscillations. In the oscillation, the inner ring was oscillating along the rotation axis while the outer ring was fixed at rotation direction, and the oscillation speed was constant. When calculating the wear of joints under the normal working condition, the relationship between the wear intensity and wear time could be considered as linear [6] . That\u2019s to say, the wear speed was a constant. The wear rate in normal running condition was decided by the surface pressure p and the sliding speed v as \u03b3 mkp v , Where k is the factor depends on the operating condition, which includes the materials of friction pairs, the quality of slid surface and the lubrication state. m, \u03c4 are the surface pressure and sliding velocity exponential factors to wear rate. 4 Advanced Research on Applied Mechanics, Mechatronics and Intelligent System The line wear rate of abrasive wear is defined as 1* mdh dh dt kp v v ds dt ds    (1) The experiments proved that the line wear rate of abrasive wear was independent to the sliding velocity [7] . So \u03c4=1, then \u03b3 mkp v (2) As shown in Fig.1(a) (the section in figure was defined by point G and the axial lead), the location of point G can be determined by the coordinates (\u03b8, a ,R). To calculate the joint wear at point G, we took a slice by the length of dl as the study object, which could be regarded as a cylindrical friction pair by the length of dl. The axial section of the slice was shown as Fig.1(b). According to JWCM, the direction of joint wear h1-2 of the so called cylindrical friction pair by the length of dl was along the x-x direction according to the working condition above, and the wear directions of h1 ,h2 were both along the normal line of the sliding surface as shown in Fig.1(b). According to Fig.1(a), the relative sliding velocity at point G can be achieved as 2 2lv nR nRsin    (3) The wear velocities at point G in outer ring and inner ring can be defined as 1 1 12m mk p v nRsin k p    , 2 2 22m mk p v nRsin k p    (4) Then the joint wear rate \u03b31-2 can be defined as 1 2 1 2 1 2( ) 2 mcos nR k k p sin cosa         ( ) (5) Then 1 1 2 1 2[2 ] mp cosa nR k k sin   \uff08 \uff08 \uff09 \uff09 (6) While from Fig.1(a), ( cos ) sindl d R R d     , sin sinldS R da dl R da R d       (7) Then   0 0 0 0 0 2 1 2 0 1 2 1 1 1 2 1 2 2 1 1 2 1 1 11 2 1 2 2 2 2 1 2 cos [ sin sin ( 2 ] ( ) ) 2 a r a a m a am m m m m a P p adS p cosa cosa nR k k sin c R dad R dad dad osa R sin cosa n k k                                           ( \uff08 \uff09 ) ( \uff09 (8) The joint wear rate \u03b31-2 can be achieved as 2 1 1 2 1 2 ( )2 m m r rn k k P R I     \uff08 \uff09 (9) Where 02 1 0 2 1 1 1(( ( ) ) ) a m m m r a I sin cosa dad          (10) \u03b81, \u03b82 and \u03b10 are the configuration parameters of ACSPB. According to Fig.1(a), we can get: 1 2arccos(( ) ), arccos(( ) )T s C R T s R      . \u03b10 is the contact half wrap angle of the so called Applied Mechanics and Materials Vol. 344 5 cylindrical friction pair. It depended on the load and mechanical property of the materials of the ACSPB, and can be got through contact mechanics method [8] . To the ACSPB which had heavy load and tiny radial clearance, \u03b10\u2248\u03c0/2. And further, from Eq. (9), \u03b31-2 can be expressed as 1 2 /r r mnKP f   (11) Where 2 1 2m r rRf I  , is the factor depended on the geometry and contact condition of ACSPB, Ir is obtained from Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003721_gt2012-68476-Figure32-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003721_gt2012-68476-Figure32-1.png",
        "caption": "Figure 32 \u2013 Schematic Concept of LE Coupon Repair",
        "texts": [
            "asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Leading Edge Coupon Weld Repair via GTAW/GMAW As observed in figure 7, there are a few occasions when the LE burns away as a result of oxidation. Out of the first 8 engine sets repaired, by coincidence the first 2 sets of production repairs exhibited this distress. Since the damage is extensive, rather than scrapping out the vanes, a LE coupon weld repair was developed, as schematically shown in figure 32. Figure 33 reveals the area of the LE that was machined away and figure 34 shows the actual GMAW process been performed to attach the new coupon in. The root pass is first performed with the GTAW process. Initial plans were to weld the entire coupon in with the GTAW process, but later it was determined that a faster repair could be achieved with the GMAW process, and as a result, the latter process was utilized to reduce repair times, as seen in figure 34. From a technical point of view both welding processes produced good quality weld repairs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure6-1.png",
        "caption": "Figure 6. Actuator",
        "texts": [],
        "surrounding_texts": [
            "No muscle is attached to clavicle. Therefore, the articulation sternoclavicularis and the acromial joint are passive joints, and only the articulation acromioclavicularis and the scapulothoracic joint are active ones. The scapula is attached to the sternum via the clavicle. Therefore the scapula motion is restricted to the spherical surface whose radius is the clavicle. Skin also restricts the scapula motion so that the scapula is always on the ribs.\nFig.3(a) shows a schematic model of these relations, and Fig.3(b) presents our 4DoF scapula model. In this model, the scapula is approximated as a rectangular board, the clavicle and the rib are approximated as links which have 3DoF joints at both ends. The 4DoF model has four sliding actuators.\nFig. 4(a) is the back view of the human body. Red areas are soft areas to which large force cannot be applied from outside. White areas in Fig. 4(a) are hard areas to which outside force can be applied. The only hard areas around the shoulder are around the articulatio acromioclavicularis and around the scapulothoracic joint. These are the points to which the trainer will apply guidance force to let the trainee learn the proper motion of the scapula and the shoulder joints. Fig. 4(b) shows corresponding points in the 4DoF model.\nIII. SCAPULA MOTION AND SCAPULA MOTION ANALYSIS\nA. Aanlysis Method by the 4DoF Model\nIn Fig. 5, \u2211 0 is the coordinate system on the human body,\nand \u2211 p is the coordinate system on the scapula ( 4321 CCCC )\nwith its origin at p O . Position and pose of the scapula in the\n4DoF model is expressed by four angles ( 1 \u03b8 , 2 \u03b8 , 3 \u03b8 , 4 \u03b8 ). 1 R ,\n2 R , 3 R , 4 R represent length variables of four sliding actuators.\nThe angle 1 \u03b8 is the angle of 113 BAA . The angle 2 \u03b8 is the angle between the triangle 113 BAA and the XY plane of \u2211 0 . The angle 3 \u03b8 is the angle between the triangle 113 BAA and the\ntriangle 331 BAB . The angle 4 \u03b8 is the rotation angle of the quadrilateral 4321 CCCC whose rotational axis is 31 BB XY . 1 a ,\n2 a , 1 b , 2 b , 1 l (length of 11 BA ) and 2 l (length of 33 BA ) are constants which are obtained by measuring the human body. 1 \u03b8\ncorresponds to elevation and depression of the scapula. 2 \u03b8 corresponds to forward and backward motion. 3 \u03b8 and angle 4 \u03b8 correspond to rotation of the scapula.\nThe transformation matrix p\nT0 between the coordinate\nsystems \u2211 p and \u2211 0 is presented by 1 \u03b8 , 2 \u03b8 , 3 \u03b8 , 4 \u03b8 . Between\nthe coordinate of a vertex i C ( 1=i , 2 , 3 , 4 ) in \u2211 p and its\ncoordinate in \u2211 0 , the following relation holds:\ni\np pi CTC 00 = \uff081\uff09\nAs the transformation matrix p T0 is expressed by 1 \u03b8 , 2 \u03b8 ,\n3 \u03b8 , 4 \u03b8 , the equation (1) can also be described as:\n( ) i p i CgC P=0 \uff082\uff09\nwhere [ ]T 4321 \u03b8\u03b8\u03b8\u03b8=P .\nWith the coordinate of the origin p O of \u2211 p in \u2211 0 ,\n[ ]T pzpypxp OOOO 0000 = , and the pose of the scapula roll \u03b8 , pitch \u03b8 ,\nyaw \u03b8 , the position and pose of the scapula in \u2211 0 can be described as [ ]T yawpitchrollpzpypx OOO \u03b8\u03b8\u03b8000=Q , and\nanother expression of the equation (1) is obtained as:\n( ) i p i ChC Q=0 \uff083\uff09\nTherefore,",
            "The force from the other actuators can be expressed in the same fashion. Then, the force balance in the 4DoF is represented by the following relations.\n( ) ( ) ( )\n \n  \n\n\u2211=\n\u2211=\n\u2211=\n=\n=\n=\n6\n1\n00\n6\n1\n00\n6\n1\n00\ncos\ncos\ncos\ni iizz\ni iiyy\ni iixx\nFF\nFF\nFF\n\u03b3\n\u03b2\n\u03b1\n\uff088\uff09\nConsidering the center of the gravity of the scapula\n[ ]T gzgygxg OOOO 0000 = , and the point [ ]T azayaxa OOOO 0000 =\nwhere the outside force F is applied , the torque balance of the 4DoF model is represented as follows.\n( ) ( ) ( ) ( )\n( ) ( )\n  \n\n    \n\n\u2211         \u2212 \u2212 =\n\u2211         \u2212 \u2212 =\n\u2211         \u2212 \u2212 =\n=\n=\n=\n6\n1\n0\n0\n6\n1\n0\n0\n6\n1\n0\n0\ncos\ncos\ncos\ni gzaz\niigziz\nz\ni gyay\niigyiy\ny\ni gxax\niigxix\nx\nOO\nFOC F\nOO\nFOC F\nOO\nFOC F\n\u03b3\n\u03b2\n\u03b1\n\uff089\uff09\nC. Scapula Motion Analysis and Force of Shoulder Mouscles\nAnalysis\nScapula motion and force of shoulder muscles are analyzed by the 4DoF model for elevation, depression, forward, and backward motions of the scapula. The force of the four actuators is regarded as the force from the muscles.\nIn this analysis, it is assumed that the human is standing upright in the natural posture, and the arm hangs from the",
            "shoulder without exercising any tasks. Therefore, the input force F to the scapula is the weight of the arm.\nConstants in the 4DoF model (Fig. 3) need to be measured in advance from the individual trainee. In this paper, average values of Japanese are used [7]. In order to calculate the center of the gravity in the 4DoF model, the weight ratio of various parts of the human bones is also necessary [7]. Table I and Table II summarize these constants, the input force F , and the weight ratios.\nFour actuators\u2019 force was calculated for a couple of simple scapula motions by the method described in the previous section. Range of scapula motion is determined from Range of Motion (ROM) [8]. When the man stands upright with his arm hanging naturally,\n1 \u03b8 , 2 \u03b8 , 3 \u03b8 , and 4 \u03b8 are 90 degrees, 90\ndegrees, 6 degrees, and 0 degree respectively. If the scapula is elevated or depressed,\n1 \u03b8 , 2 \u03b8 , 3 \u03b8 , and 4 \u03b8 change in the ranges\nshown in Table III. The scapula at 70 1 =\u03b8 degrees corresponds to the posture where the shoulder is moved to the uppermost position. When the shoulder is at the lowermost position,\n1 \u03b8 is\n110 degrees. Likewise, if the shoulder is moved from the backward limit position to the forward limit position,\n2 \u03b8\nchanges from 70 degrees through 110 degrees (Table IV).\nFig.8 shows scapula\u2019s position p O pose roll \u03b8 , pitch \u03b8 , yaw \u03b8 , and\nFig.9 shows the forces from the four actuators, when the shoulder is moved in the range shown in Table III. Fig.10 and 11 show the position, pose, and forces calculated for the forward and backward motion in Table IV."
        ]
    },
    {
        "image_filename": "designv11_100_0002780_s10851-011-0286-y-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002780_s10851-011-0286-y-Figure5-1.png",
        "caption": "Fig. 5 Leg geometry. Transformations applied to determine the leg tip coordinates",
        "texts": [
            " In order to determine which leg tips are the supporting points, we proceed as follows: (1) We compute the tip position of each leg relative to the body center coordinate reference system from the knowledge of the joint angles of each leg articulation. (2) We compute the hypothetical supporting planes defined by each combination of three leg tip points. (3) We discard hypothetical supporting planes for which at least one leg tip is below it or do not comply with the stability condition about the center of mass. First we compute the position of the leg tip g applying the transformations defined by the chain of articulations from the leg tip up to the body center, as shown in general in Fig. 5. Figure 6 shows the leg parameter specification for the Aibo robot. The In homogeneous coordinates, the leg tip is computed by the following product of elemental transfor- mation matrices: ( g 1 ) = (Tn+1 \u00b7 Rn.Tn \u00b7 \u00b7 \u00b7R1.T1). ( 0 1 ) , (13) where Rk is the rotation matrix corresponding to the k-th leg articulation from tip to the body center, being n the number of articulations and Tk\u22121 the translation matrix corresponding to the leg segment between the (k \u2212 1)-th and the k-th articulations. Translation matrix T1 corresponds to the translation from the tip to the first articulation, while translation matrix Tn+1 corresponds to the translation from the last articulation to the body center reference system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.108-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.108-1.png",
        "caption": "Fig. 2.108 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RPaRRR (a) and 4PaRRPR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology R\\Pa\\kR\\R||R (a) and Pa||R\\R\\P\\kR (b)",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.65-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.65-1.png",
        "caption": "Fig. 2.65 4PaPaRP-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\Pa\\\\R||P",
        "texts": [
            "22q) Idem No. 33 45. 4PPaPaR (Fig. 2.57) P||Pa||Pa\\R (Fig. 2.22r) Idem No. 4 46. 4PaPPaR (Fig. 2.58) Pa||P||Pa\\R (Fig. 2.22s) Idem No. 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002326_tmag.2013.2247984-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002326_tmag.2013.2247984-Figure2-1.png",
        "caption": "Fig. 2. Displacements of the rotor at two bearings: Type-1 response (left) and Type-2 response (right).",
        "texts": [
            " Often, the coefficients from two bearings are in the same order and seldom differ by more than two orders of magnitude. For the ease of explaining the response, we will assume that this condition is met for the rest of this section. Type 1: . This occurs when the inertia force of the rotating part is much smaller than the damping and restoring force of the upper bearing. As a result, (8) implies (12) There are several features in this type of response. First, the most important feature is that the displacements at the two bearings are out of phase (Fig. 2), as signified by the negative sign in (12). The out-of-phase relationship is necessary to maintain the rotating part in dynamic equilibrium in the radial direction, because the inertia force is now negligible. Second, the ratio of the complex displacement at the two bearings is inversely proportional to the impedance of the corresponding bearing forces. Third, and are out of phase and roughly the same order; therefore, their difference will be of the same order. So none of them can be ignored if (10) or (11) is used to extract and "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002125_978-3-642-23147-6_34-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002125_978-3-642-23147-6_34-Figure5-1.png",
        "caption": "Fig. 5. The illustration of the position for each robot, obstacle and target ball",
        "texts": [
            " Step 5: We set the kinematic value or propotional gain, \u039a based on . Step 6: We calculate the velocity of the robot to move left, or right, as below. \u039a (8) \u039a (9) where = velocity control of the robot heading direction. Step 7: End. In addition, the proposed passing flow chart is depicted as shown in Fig. 4. After R1 has passed the ball to R2 using the above passing strategy, we construct another strategy for moving with obstacle avoidance and shooting the ball. Based on the current positions of the robot, obstacle and the target ball (Fig. 5), we predict the robot velocity and directions to the target position. Let us say, if the opponent robot (obstacle , ) exists in between the current , , the desired , , and the target positions , , the robot must avoid the obstacle if the perpendicular distance to the line, l of the obstacle, d is less than radius, ro. The algorithm (Fig. 6) for shooting with or without obstacle avoidance is explained as algorithm 2 [12]. Algorithm 2. The proposed moving with obstacle avoidance and shooting algorithm Input: Current position of all the robots and the ball, and the subsequent position of the robots"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002563_ijmms.2013.052783-Figure13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002563_ijmms.2013.052783-Figure13-1.png",
        "caption": "Figure 13 Deadlock (all the robots are continuing to avoid a collision)",
        "texts": [
            " It is seen from these results that there is a large discrepancy in time consumption between two cases, depending on considering the collision avoidance or not. Note however that the time required to converge is not exponential with respect to the number of robots, but proportionate. Figure 12 shows the success rate [%] in forming a desired formation. From this figure, observe that the success rate drastically decreases, if the number of robots exceeds seven. This is attributed to the fact that the method for collision avoidance is incomplete. There may happen a deadlock situation as shown in Figure 13. When every robot wants to go to own destination, they have to turn to the left 90 [deg] to avoid collision, because there are other robots near the route. Thus, if all the robots perform the same thing simultaneously, any robots cannot reach own destination forever. To overcome this situation, it was assumed that the robot speed was changed randomly [0%, 100%] within a given range at random time [0 sec, 1 sec]. Thus, the improvement of such a method is required to increase the success rate of forming a formation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.5-1.png",
        "caption": "Fig. 2.5 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRPR (a) and 4RRPRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R||R\\R\\P\\||R (a) and R||R||P\\R||R (b)",
        "texts": [
            "1b) The first, the second and the last revolute joints of the four limbs have parallel axes 3. 4RRRRR (Fig. 2.3a) R||R||R\\R||R (Fig. 2.1c) The two last revolute joints of the four limbs have parallel axes 4. 4RRRRR (Fig. 2.3b) R||R||R\\R||R (Fig. 2.1c) The three first revolute joints of the four limbs have parallel axes 5. 4RRRRR (Fig. 2.4a) R||R\\R||R||R (Fig. 2.1d) The two first revolute joints of the four limbs have parallel axes. 6. 4PRRRR (Fig. 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7. 4RRRPR (Fig. 2.5a) R||R\\R\\P\\kR (Fig. 2.1f) Idem No. 5 8. 4RRPRR (Fig. 2.5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10. 4PRRRR (Fig. 2.6b) P||R\\R||R\\R (Fig. 2.1i) Idem No. 9 11. 4RRPRR (Fig. 2.7a) R\\R\\P\\kR\\R (Fig. 2.1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2.8a) P||R||R||R||R\\R (Fig. 2.1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003946_amm.120.343-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003946_amm.120.343-Figure1-1.png",
        "caption": "Fig. 1 Load distribution of the wire race ball bearing under the radial force",
        "texts": [
            " The other is the full finite element analysis [7]. However, the full finite element contact model is greatly complex, for taking all contact elements into account. And the solving process highly cost computing resources. Consequently, the prior method is preferred in this work. The Stribeck theory is defined by that: the loading zone is a half of the circumference when a radial load is applied on the bearing; and the resultant forces from all balls which withstand different loads maintain a balance with the radial load. Fig. 1 illustrates the load distribution of the wire race ball bearing under the radial force. From the force equilibrium illustrated in Fig. 1, the mathematical description of the radial load in a wire ball bearing can be presented as follows 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 TTP, www.ttp.net. (ID: 132.174.255.116, University of Pittsburgh, Pittsburgh, USA-18/03/15,06:12:09) ( ) 4 1 2 2 cos 1 Z Fix r B Bj j F F F j \u03c8       = = + \u2212  \u2211 . (1) where Fr is the radial load acting on a wire race ball bearing (N), FB1 is the contact load acting on the ball labeled B1 which has the maximum load (N), \u03c8 is the position distribution angle of the rolling element (rad), and 2 Z \u03c0 \u03c8 = ; 4 Z Fix       denotes an integer resulted from rounding the fraction 4 Z toward zero, j is an integral number, and 1 4 Z j Fix  \u2264 \u2264     (Z denotes the number of balls)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001340_b978-081551497-8.50005-6-Figure3.28-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001340_b978-081551497-8.50005-6-Figure3.28-1.png",
        "caption": "Figure 3.28 Shape optimization redistributes the boundary of a fixed topology to find an optimal shape. Adapted from Haftka and G\u00fcrdal.[34]",
        "texts": [
            " It is clear that topology optimization is the highest level of structural optimization with which to optimize the topology, shape, and size simultaneously. Some details of shape and topology optimization are given below. Shape Optimization. One key point of shape optimization is to choose a method with which to represent the boundary of the design domain, because the optimal shape depends on design variables selected to represent modifiable boundaries. For example, the optimal shape of the cantilever beam shown in Fig. 3.28 depends on the boundary representation and the number of design variables. If the cantilever beam is modeled as shown in Fig. 3.28(a), with only one design variable (the height H of the free end), then the optimal shape will be different from the model shown in Fig. 3.28(b), where the boundary is represented by a cubic spline with control points along the length of the beam. The most commonly used methods to represent a boundary are based on polynomials and splines. After the boundary representation is complete, the analysis module passes the geometrical model to the analysis model through an analytical or numerical method, because the geometrical model cannot be used directly in most cases. The FEM and BEM are generally applicable techniques. For example, FEM could generate a mesh inside the design domain automatically; adaptive remeshing or mesh refinement would fol- MEMS AND NEMS SIMULATION, KORVINK, RUDNYI, GREINER, LIU 143 CH03 9/9/05 8:50 AM Page 143 low the boundary movement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.2-1.png",
        "caption": "Figure 10.2 The flux distributions of a four-pole induction motor during transient finite element analysis",
        "texts": [
            " In other words, the total magnetic field is a sinusoidal field with its peak rotating at angular speed \u03c9 rad/s. Since \u03c9 = 2\u03c0f , the rotating speed of the field will be the same as the supply frequency: f revolutions per second or nS = 60f revolutions per minute (rpm). Noting that the above derivation is based on one pair of poles, a more generic equation for the field speed (or synchronous speed) of an induction machine can be given as nS = 60f p and \u03c9S = 2\u03c0nS 60 = 2\u03c0f p = \u03c9 p (10.4) where p is the number of pairs of poles. Figure 10.2 shows the arrangement of a four-pole squirrel-cage induction motor with flux distribution. Assuming initially that the rotor is stationary, an electromotive force (emf) will be induced inside the rotor bars of the squirrel cage. A current is therefore formed inside the rotor bars through the end rings. Similarly, since the field is rotating, this current will generate a force on the rotor bars (the rotor bar current is under the stator magnetic field). If the force (or torque) is sufficiently large, the rotor will start to rotate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.116-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.116-1.png",
        "caption": "Fig. 2.116 4PaRRRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 26, limb topology Pa||R||R\\R||Pa",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002965_amr.328-330.743-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002965_amr.328-330.743-Figure1-1.png",
        "caption": "Fig. 1 Movements of Taking Box",
        "texts": [
            " It takes drug box out from shelf and shove it off during the process of operation, at last, put the box on the conveyor belt smoothly in order that the medicine board can be put into it in next station [4]. On the whole, box-taking mechanism is a \u201cbottleneck\u201d in the cartoner. Its efficiency affects the whole efficiency. Therefore, developing a high quality box-taking mechanism will be a sally port to solve the problem of the high-speed cartoner. In order to improve the efficiency of taking box, the present box-taking Mechanisms are using continuous rotary structure. Movements of taking box request actuators have a complicated trajectory. As shown in Fig. 1. Rotating vacuum suction head 1 suctioned the box out from the shelf 2, utilizing extrusion between medicine box and block 5 to shove box off slightly, then put it on the conveyor belt 4. The entire movements can be summed up as: suctioned box-shove box-put box down. The most important movement is the process of box-suctioning. When the vacuum suction head contacts box surface vertically, there must be a process of mutual extruding, making the 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_100_0001858_978-94-007-7401-8_2-Figure2.110-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.110-1.png",
        "caption": "Fig. 2.110 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RPRRPa (a) and 4PRPaRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology R\\P\\kR\\R||Pa (a) and P\\R\\Pa\\kR\\R (b)",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003161_amm.105-107.244-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003161_amm.105-107.244-Figure2-1.png",
        "caption": "Figure 2 Boundary conditions as substitution of lifting devices.",
        "texts": [
            " Before the model properties were identified the mechanism was simplified in order to reduce the hardware and time demands [2]. The stiffness of lifting devices was determined via FEM (Finite Element Method) and consequently compared with stiffness of other components (tables, pillars). Since the stiffness of both lifting devices is of higher order, their modeling can be realized by appropriate boundary condition \u2013 by removing of degrees of freedom in vertical direction in the location of connection of lifting devices and tables \u2013 see Fig. 2. The base frame firmly coupled with the ground was substituted by appropriate boundary condition as well. Its influence on the modal properties can be neglected. Regarding to this fact all degrees of freedom of the pillars were removed in the location of their connection with the base frame. The beam elements (pillars) [3] and shell elements (tables, reinforcements) were used in the finite element model. The device was loaded by the mass element located in the middle of both tables \u2013 symmetrical loading of 1000kg was assumed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002396_ssp.198.324-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002396_ssp.198.324-Figure1-1.png",
        "caption": "Fig. 1. The scheme of system of several pneumatic actuators: 1 \u2013 oscillatory mass, 2, 3 \u2013 pockets of actuators, 11",
        "texts": [
            " It has been determined that at certain mechanical and dynamic parameters of the system, when selfexciting vibrations are initiated in one of the vibrodrives, the system also enters the self-exciting vibration mode. The effect of the vibrodrive geometric parameters on the changes in the law of motion of the system working organ has been studied. Theoretical study of system of pneumatic actuators A scheme of the investigated system of several pneumatic actuators connected by a rigid link (in this case, a rigid link is total oscillatory mass of the system of several actuators) is presented in Fig. 1. The motion of the working links of actuators 1, the change of the gas mass in the chambers 2, 3 and the working clearances 4, 5 of the actuators, and gas inflow into the chambers 2, 3 and its outflow from its, is defined by the equations systems, presented in [1, 2]. According to the worked out mathematical model, the software was used for studying the total laws of motion of the working links of several pneumatic actuators. The study results are presented in Fig. 2 \u2013 Fig. 4. The area in which synchronic operation of several pneumatic actuators connected by a rigid link exists presented in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002826_s1644-9665(12)60146-0-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002826_s1644-9665(12)60146-0-Figure3-1.png",
        "caption": "Fig. 3. The constructional solution for the supported column structure",
        "texts": [
            " These are: jointed (both column ends were fixed to rotating loops) and supported (one column end fixed to a rotating loop and the other strengthen). The constructional solution for the jointed Euler loaded column is shown in Figure 2. The active load charging head is the prism 1(1). The head taking the active or passive column 4 load is the set of shaft 2 and accordingly: needle rolling bearing 3(1) \u2013 column A \u2013 or stiff cylindrical element 3(2) \u2013 column B. The head taking the load \u2013 prism 1(2) \u2013 is fixed to base 5. The constructional solution for the supported Euler loaded column is shown in Figure 3. The difference between this solution and the column in Figure 2 is the direct fixing of one end of the column 4 onto base 5. The columns are labelled according to the cylindrical element in use: \u2013 column C \u2013 rolling ball bearing 5(1), \u2013 column D \u2013 stiff cylindrical element 5(2) with circular contour of the working area. A loss of stability occurs in the plane with lower bending stiffness, after exceeding a certain value of the axial force P called critical force Pc. The formulation of the vibration topic is based on the Hamilton rule using the Bernoullie-Euler theory"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001959_9781119969242.ch7-Figure7.26-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001959_9781119969242.ch7-Figure7.26-1.png",
        "caption": "Figure 7.26 Non-inverting Schmitt trigger and wave shaping circuit",
        "texts": [],
        "surrounding_texts": [
            "The inverters control logic can be implemented using an analog circuit and its complete block diagram is shown in Figure 7.24. The other method is by using advanced microprocessors, microcontrollers, and digital signal processor to implement the control schemes. In Figure 7.24, the supply is taken from a single-phase power supply and converted to 9-0-9Vusing a small transformer. This is fed to the phase shifting circuit, shown in Figure 7.25, to provide an appropriate phase shift for operation at various conduction angles. The phase shifted signal is then fed to the inverting/non-inverting Schmitt trigger circuit and waveshaping circuit (Figures 7.26 and 7.27). The processed signal is then fed to the isolation and driver circuit shown in Figure 7.28, which is then finally given to the gate of IGBTs. There are two separate circuits for upper and lower legs of the inverter. The power circuit can be made up of IGBT, with a snubber circuit consisting of a series combination of a resistance and a capacitor with a diode in parallel with the resistance."
        ]
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.2-1.png",
        "caption": "Fig. 3.2",
        "texts": [
            " Thus, the motion of a single point of the body arbitrarily chosen represents the motion of the complete body. During a rotation all the particles of a rigid body move about a common axis. In the special case that this axis is fixed in space, the motion is called a rotation about a fixed axis. If on the other hand the axis only passes through a fixed point without keeping its direction, then the motion is referred to as a rotation about a fixed point or a gyroscopic motion. Let us first consider the motion of a rigid body about a fixed axis (Fig. 3.2). In this case each point of the body moves in a circle whose plane is perpendicular to the axis. The radius vectors from the axis to the individual points of the body sweep out the same angle d\u03d5 during the same time interval dt. Thus, the angular velocity \u03c9 = \u03d5\u0307 and the angular acceleration \u03c9\u0307 = \u03d5\u0308, respectively, are the same for every point. The velocity and the acceleration of an arbitrary point P at a distance r from the axis are therefore the same as for a particle in a circular motion (see (1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.114-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.114-1.png",
        "caption": "Fig. 2.114 4PaRPaRR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 26, limb topology Pa||R\\Pa\\kR\\R",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002280_synasc.2013.9-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002280_synasc.2013.9-Figure4-1.png",
        "caption": "Figure 4. Tape after two folds",
        "texts": [
            " We take arbitral points E and F and an arbitrary rectangular origami, and further assume, without loss of generality, that the height of the origami is of unit length, and compute Gro\u0308bner bases of the set of the polynomials generated from the logical specification of the propositions, over the coefficient domain of rational functions. We will discuss more in detail in Section V. After the fold that constructs the isosceles triangle \u25b3 GFE, we perform another fold to construct another isosceles triangle in such a way that both isosceles triangles \u25b3 GFE and \u25b3 FGH overlap as shown in Fig. 4. Lemma 2 states the construction involved and the resulting geometrical property. We define the predicate O5(P, \ud835\udc60,Q, \ud835\udc65) to express the relationship among the point P and Q, and the lines \ud835\udc60 and \ud835\udc65, which are involved in Huzita\u2019s basic fold (O5). It states that the fold line \ud835\udc65 passing through point Q superposes point P and line \ud835\udc60. Lemma 2 (Isosceles trapezoid). For any origami ABCD, point E \u2208 AB, point F \u2208 CD and point G \u2208 CD such that G \u2208 EAEF, for any line \ud835\udc5b and points H and I such that O5(F, EB, G, \ud835\udc5b) \u2227 H \u2208 AB \u2227 F \u2208 HB\ud835\udc5b \u2227 I \u2208 FDEF \u2227 I \u2208 GC\ud835\udc5b, we have \u2223EF\u2223 = \u2223GH\u2223. Figure 4 shows the constructed tape that will help us to see the geometrical meaning of the lemma. The lemmas that we treat by the EOS prover have the following structure: For any \ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc56 such that \ud835\udcab(\ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc56), for any \ud835\udc5c\ud835\udc56+1, . . . , \ud835\udc5c\ud835\udc58 such that \ud835\udcac(\ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc58) some geometrical properties hold among the objects \ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc58, where (i) \u201cFor any \ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc56 such that \ud835\udcab(\ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc56)\u201d is the configuration before the construction. (ii) \u201cfor any \ud835\udc5c\ud835\udc56+1, . . . , \ud835\udc5c\ud835\udc58 such that \ud835\udcac(\ud835\udc5c1, . . . , \ud835\udc5c\ud835\udc58)\u201d is the step of the construction (iii) \u201csome geometrical properties hold among the objects \ud835\udc5c1, ",
            " If we omitted the description about the point I, which will be needed in Theorem 1 as well, we can have other two possible fold lines. The fold along those lines falsify the goal \u2223EF\u2223 = \u2223GH\u2223. We tend to overlook these kinds of geometrical phenomena unless we perform the proof by a computer-assisted prover. By overlapping the two crossings we construct the isosceles trapezoid, which is a basis for constructing regular polygons. Now a yet one more fold creates a pentagon-like shape. Namely from the tape of Fig. 4, we perform a fold along a fold line that passes though F, to superpose H and CG. Note that this is realized by Huzita\u2019s basic fold (O5). There are two possible fold lines, say \ud835\udc5a1 and \ud835\udc5a2 to make this possible. The fold along \ud835\udc5a1 creates the shape as shown in Fig. 5. The other case of the fold reflects the point I across \ud835\udc5a2 and does not make EHGIF a pentagon. Finally, to make a knot, we have to perform a valley fold and insert the moving face of the tape in between the existing non-moving faces of the tape, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.49-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.49-1.png",
        "caption": "Figure 3.49 Coordinate frames for rapid rotation of a fighter aircraft",
        "texts": [
            " Furthermore, the instantaneous direction of the velocity vector, v, can be regarded as fixed during the rapid attitude maneuver, which implies that the changes in the angle of attack, \u03b1, and sideslip angle, \u03b2, are only due to the change in the aircraft attitude. Consequently, during a rapid aircraft rotation one can take the stability axes, (ie, je, ke), to be essentially aligned with the instantaneously fixed wind axes, (iv, jv, kv), and used as a reference frame for measuring the aircraft attitude, (\u03b1, \u03b2, \u03c3), as shown in Figure 3.49. However, now that the displaced attitude can be large, we refrain from using the traditional Euler angles that can have singularities during a large maneuver. Instead, we employ a non-singular attitude representation given by the attitude parameters vector, \u03b6(t), such as the quaternion or the modified Rodrigues parameters (Tewari 2006). The angular velocity, \u03c9, and external torque, \u03c4, vectors are resolved in the instantaneous body axes, (i, j, k), as follows: \u03c9 = P i + Qj + Rk \u03c4 = Li + Mj + Nk"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002003_978-3-642-27329-2_25-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002003_978-3-642-27329-2_25-Figure3-1.png",
        "caption": "Fig. 3. Force analysis of the moving mass system",
        "texts": [
            " The control target of anti-rolling device is to keep the ship stable by viewing the wave moment as noise and reducing the noise: ( ) ( ) ( )2xx mI C N B Dh A Dh\u03c6 \u03c6 \u03c6 \u03b1\u2032 + + + + + = Fig. 2 shows the 3d model of the moving mass system. In this system[5], the rail of the moving mass is placed across the deck, and the mass is moving along the transverse direction of ship. The distance between the moving mass and ship gravity center is Zm. So the 3d model can be simplified as 2d model just considering the single-degree-of-freedom roll. Fig. 3 shows the 2d force analysis of the moving mass system. is the position of the mass; F is the force of actuator exerting on the mass; Mc is the control moment; is roll angle; a is the acceleration of the moving mass. From the analysis of force balance on the mass along the transverse direction of ship, . (7) . (8) The reaction of the mass generates the control moment Mc, which consists of moments coming from reaction F (relative to F) and N (relative to normal component of gravity mgcos ). . (9) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002287_nano.2011.6144391-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002287_nano.2011.6144391-Figure2-1.png",
        "caption": "Figure 2: Schematic illustration of experimental rack.",
        "texts": [
            " In order to simulate the microgravity condition, the plane follows a parabolic trajectory starting from a normal flight level in about 7925m, being pitched up to about 8535m with an inclination angle of about 45\u02da at the beginning of the parabolic. The plane is being guided into the free fall part of the trajectory, where thrust compensates the remaining air drag, creating a microgravity condition. The gravity will be recovered after the end of the free fall period as shown in Figure 1. The volumes of the ferrofluid in this experimental setup are 0.1ml and 0.5ml. The polarity of the magnet is also being manipulated to observe the difference between them. Setup of the sample kit is shown in Table II. An experimental rack was setup [Figure 2], which mainly consist of one sample kit and two camcorders. Cam T gives the top view and Cam T gives the side view observations. The sample kit is used to place the ferrofluid. Both of the camcorders are left turned on throughout the whole parabolic flight. 978-1-4577-1515-0/11/$26.00 \u00a92011 IEEE 1030 The sample kit consists of two layers with four identical cells each, with a magnet holder consist of four magnets placed between the layers. [Figure 3] The magnet is adjusted to be in the centre of each cell of the sample kit, which prevents the interaction of the magnet with the ferrofluid of another cell"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002459_urai.2012.6463047-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002459_urai.2012.6463047-Figure3-1.png",
        "caption": "Fig. 3. Reader antenna Moved type RFID tag reading system",
        "texts": [
            "1 Reader Antenna Fixed type Tag Reading System In the reader antenna fixed type RFID tag reading system, reader antennas are arranged in a matrix pattern under the workpiece tray (Fig. 2.). These fixed antennas communicate the tags attached to the workpiece, and the workpiece position and posture are measured. This system can read the data in the tags instantly after the workpieces are put on the tray. 2.2 Reader Antenna Moved type Tag Reading System In the reader antenna moved type RFID tag reading system, the data in tags are read by relative motion of the RFID tag and the tag reader (Fig. 3.). Workpiece position and posture are measured by the start and stop time of tag reading and the speed of the reader movement. This system can reduce the number of tag readers compared with the fixed type system. Measurement errors were analyzed on the following conditions. Two RFID tags T1 and T2 are attached to a workpiece, and the communication range of the reader is arranged in a matrix pattern. The communication range has a width of a in the direction of x-axis and a length b in the direction of y-axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001884_9781118609811.ch17-Figure17.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001884_9781118609811.ch17-Figure17.2-1.png",
        "caption": "Figure 17.2. Schematic fabrication of mesoporous silica nanoparticles. (See color insert.)",
        "texts": [
            "27 Porous Silica Nanostructures. Notwithstanding minimum diffusion limitations, enzyme loading per unit mass of nonporous nanoparticles is usually low. On the other hand, porous nanoparticles can afford high enzyme loadings due to their large surface areas. It should not come as a surprise that mesoporous silica particles have attracted a significant amount of attention as matrices for enzyme immobilization due to their high surface areas, controllable pore diameters, and uniform pore size distributions (Figure 17.2).28\u201330 Numerousmesoporous silicas have been synthesized and utilized for enzyme immobilization applications. The list of mesoporous materials used for this purpose includes MCM-41, SBA-15, and mesocellular foam (MCM).28,31, 32 Many studies have reported improved enzyme stability after their immobilization in mesoporous silicas.33\u201336 In addition, the immobilization behavior of enzymes in mesoporous silicas was found to depend on the silicamorphology37 and the textural and structural parameters.38 In their study, the authors investigated the effects of the matrix structure (cubic or hexagonal), the nature of the pores (channel-like or cage-like), the connectivity of the porous network, and the pore size"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003240_amr.591-593.96-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003240_amr.591-593.96-Figure6-1.png",
        "caption": "Figure 6 ANSYS model of carrying frame and stress distribution",
        "texts": [
            " To set two analysis paths in the carrying frame, path1 runs through the centre part of carrying frame, path2 takes up along the sideline of carrying frame. The maps of the two paths are shown in Figure 3. According to these two paths values are analyzed from bottom to top for the carrying frame. Figure 4(a) and Figure 5(a) show the deformation curves of horizontal deformations, vertical deformations and the overall deformations along path 1, 2. Figure 4(b) and Figure 5(b) show the stress distribution curves along the path 1 and 2. Figure 4(c) and Figure 6(c) show geometry distribution of stress along the path 1, 2 [9] . The stress distribution of carrying frame is showed in Figure 3(b). With bending in any angle, the maximum stress of the overall brace concentrated on the outside of the location lower backrest was fully constrained, i.e. the maximum moment position. Due to the free freedom of the upper carrying frame, the backrest stress distribution is showed in Figure 3(b), and the maximum stress range is 32MPa~174MPa in the process of entire bending",
            " But the situation body back fully fit with backrest still cannot be met. Verified by experiments, in the process of bending over, the human body and the backrest will be separated necessarily starting from the bottom of the backrest 2/3 position. Therefore the degree of freedom perpendicular to the direction of carrying frame in the lower part of backrest 2\\3 position will be restricted to simulate the actual fit situation of the human body and the backrest on the condition with human support. Figure 6(a) shows the ANSYS model of backrest structure, the model constraints and load distribution maps on the human body support condition, and 10 load steps are set for multi-load step analysis on the condition of human body support. The stress distribution of carrying frame is showed in Figure 6(b). With bending in any angle, the maximum stress of the overall brace concentrated at the critical position that the body has not been fit. Figure 7 and Figure 8 show the backrest horizontal deformation and vertical backrest direction deformations are basically the same from bottom to up, and the stress of the middle backrest is concentrated on upper the 1/3 part of the backrest, while the stress of upper and lower 1/3 part of carrying frame on the side of backrest is large and the middle is small"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure12-1.png",
        "caption": "Figure 12. Suspension Module Analysis.",
        "texts": [
            " Engineering plastic injection gate is 3-Side, injection pressure is approximately 40MPa, injection time is 1.7sec, plastic melting temperature is approximately 280 degree and mold temperature is approximately 80 degrees was developed by setting and mold as shown in Figure 11. In the upper arm CAE analysis, the result of composite material stiffness is more than the equivalent level compared to forged steel. Front suspension module CAE modeling applying composite material suspension is made as shown Figure 12. So lateral force happened to vehicle apply to CAE modeling and lateral stiffness analysis of front suspension module has been implemented. The result of composite material suspension module of lateral stiffness is more than the equivalent level compared to Forged steel suspension module as shown in Figure 13. Ball joint that deliver from the tires force is an important part to higher strength and durability. And in Hyundai Motors is based on ball joint test standard, durability test was performed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001504_1-4020-3169-6_48-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001504_1-4020-3169-6_48-Figure1-1.png",
        "caption": "Fig. 1. Prediction of the magnetic noise in a DC electric motor",
        "texts": [
            " A special emphasis is given to the investigation of different loading conditions and two design variations of the rotor skewing angle. One motor had a rotor with no skewing and the other had a rotor with a skewing of one rotor slot. The numerical results obtained with the FEM and/or the BEM are compared with the experimental data. \u00a9 2005 Springer. Printed in Great Britain. Magnetic noise can be calculated from a three-times sequentially coupled electromagneticmechanical-acoustic numerical model [1\u20133], [6]. Fig. 1 shows a typical model of this kind. To calculate the magnetic forces that excite the structure of the motor, an electromagnetic model was developed. By applying the magnetic forces to the structural model we can calculate the magnitude and frequencies of the resulting vibrations. Finally, the results of the structural analysis, represented by the velocities on the exterior surface of the electric motor, are used as an input for the BEM acoustic model. Magnetic Forces For the magnetic-force calculation in the investigated DC electric motor, two 3-D FEM models were built; the only difference between the models was in the rotor skewing angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002441_amr.544.286-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002441_amr.544.286-Figure6-1.png",
        "caption": "Fig. 6 Steering principle of present bogie Fig. 7 Dynamic balance force system of wheelset",
        "texts": [
            " In addition, macro-hardness examinations show that surface and core of the outer ring have respectively values of 47 and 29 HRC, which are lower than the required ranges of 59 to 63 and 32 to 48 HRC, respectively. These indicate that the fractured outer ring is with a poor manufacturing quality. Table 1 Chemical composition of the bearing Composition C Si Mn Cr Ni Mo P S Cu Ti Al Ca Test value 0.22 0.24 0.50 0.47 1.71 0.25 0.009 0.009 0.090 0.001 0.028 0.0002 Required value [2] 0.17-0.23 0.15-0.40 0.40-0.70 0.40-0.60 1.60-2.00 0.20-0.30 \u22640.020 \u22640.020 \u22640.20 \u22640.005 0.010-0.050 \u22640.001 Force Model. As shown in Fig. 6, on a curved railway line the outer wheel of first wheelset in the car bogie is subjected to a passive steering force Fg; companied vice-frame will then make the linking rode a to be subject to a tensile force Ft and the linking rod b to be subjected to a press force Fc; then, a steering moment MT yields for the bogie. In addition, the wheelsets are also subjected to forces from bogie side frames through the vice-frames. Dissembled the linking rods and the side frames, one wheelset, one vice-frame, and two pieces of rails forms a merged part with clear dynamic balance force system as shown as in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.15-1.png",
        "caption": "Fig. 2.15 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RPRRP (a) and 4PRRRP (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\P\\kR\\R\\P (a) and P\\R||R\\R\\P (b)",
        "texts": [
            "1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2.1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig. 2.1g0) Idem No. 13 34. 4RRRC (Fig. 2.19a) R\\R||R||C (Fig. 2.1h0) Idem No. 21 35. 4RCRR (Fig. 2.19b) R\\C||R||R (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002859_amr.785-786.1172-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002859_amr.785-786.1172-Figure6-1.png",
        "caption": "Fig. 6 Stress nephogram after optimization(0\u00b0working condition)",
        "texts": [
            "155 mm, the diameter of steel ball radius is 13.509 mm , bell-like raceway (highest spot) is evenly distributed by 47.2 mm in diameter, the star -likeoraceway (highest spot) is evenly distributed by 46.2 mm in diameter, the rotate diameter of steel ball is 46.5 mm. When this universal joints work under 0\u00b0working condition, maximum stress is 1711.3 Mpa after optimizing the raceway, the maximum stress location is still at Y-axis 0\u00b0on the steel ball, the weakest steel ball stress reduced 14%after optimization, as shown in fig. 6. When this universal joints work under 34\u00b0limiting working condition, maximum stress is 1995.5 Mpa after optimizing the raceway, the maximum stress location is at Y-axis 0\u00b0on the steel ball, the weakest steel ball stress reduced 10%after optimization, as shown in fig. 7. (1) when the engine output the biggest torque , the maximum stress appeared on the steel ball, under 0 \u00b0working condition . The maximum stress also appeared on the ball under 34\u00b0 limiting working condition. We can see the steel ball is the weakest part of the universal joint, and its fatigue life is the shortest in the process of use"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure2-1.png",
        "caption": "Fig. 2 Quick connect wall outlet for medical gas service outlet",
        "texts": [
            " This kind of mechanism can be used in a situation that a connection of a gas or fluid supply tube and a gas or fluid manifold, such as an oxygen setup. It can prevent incorrect connection of the supply tube to the required supply source. An example of this kind of connection is a quick connect wall outlet for medical gas service outlet [5].It\u2019s a connector valve for connection to a medical gas adaptor. Medical gas service outlets are common in hospital rooms, and this kind of quick connector is well used in this situation. On the basis of mechanical structure of pneumatic connector, this has been developed. Fig.2 is a sectional view of the medical gas service wall outlet. The wall outlet comprising an index plate, a closure member connected to the index plate, and a connector valve for connection to a medical gas adaptor. The valve is connected to a backside of the index plate and is accessible from the front of the outlet when the closure member is removed. The valve includes an open-ended valve housing 1 that has a first O-ring 2 placed around a distal end 3 of the assembly. A plunger spring 4 is placed within the assembly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002952_irase.4.2013.1.10-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002952_irase.4.2013.1.10-Figure1-1.png",
        "caption": "Fig. 1. Calculation of input portion. a) in plane X\u2013\u03c6, b) in plane Z\u2013\u03c6",
        "texts": [
            " In the case of the special roller transmissions the main screw surfaces consist of three parts: the acceleration (the input), the central helical screw and that the deceleration (the output). The acceleration portion is necessary for obtaining the cheeping speed of the tool upon entry into helical channel of the pieces. If this entry portion is short, than helical surface pitch errors may occur. The deceleration portion is required to allow time to reduce the speed until the change of direction of travel takes place. This is necessary to return the starting position for a new crossing. In order to calculate these curves we elaborate one specifi c method. In Fig 1, we present the method NEW TECHNOLOGY FOR ENVIRONMENTAL FRIENDLY MANUFACTURING OF THE COMPLEX HELICAL SURFACES CS. GYENGE 1a, L. OL\u00c1H 2, M. CALIN 1 1 Technical University, Cluj-Napoca, Romania 2 ROTOTCRAFT A.G. Switzerland a E-mail: Csaba.Gyenge@tcm.utcluj.ro In the last years there have been a host of new transmissions with high effi ciency and high precision. In the fi rst stage of their manufacturing technologies did not take environmental issues into account. The research by our group has proposed to develop a technology that besides providing constructive parameters of transmissions aims not to pollute the environment too much. The developed new technology is based on ultraprecision turning, and was tested on a HEMBURG ultrapecision equipment. Keywords: environmental manufacturing, gearing, precision turning 70 Int. Rev. Appl. Sci. Eng. 4, 2013 of calculation for the acceleration portion. In Fig. 1 we used the following notations: R0 \u2013 bottom half of the thread diameter of the circle, Re \u2013 entry area, H \u2013 thread pitch as Z axes, Rn = Ra \u2013 external radius of thread. max 01 max 2 2 21 01 max 01 012 ( )cos 4( ) cos 4( 2 ). nou nou nou e e e R R R R R R R R \u03d5 \u03d5 = + \u2212 + \u2212 + (2) 3. Cutting depth determination Int. Rev. Appl. Sci. Eng. 4, 2013 71 where sin 1 1 .M d a r r\u03b5 \u03d5 \u239b \u239e = \u2212 \u2212\u239c \u239f+\u239d \u23a0 (4) For manufacturing the special helical channels with the ultraprecision machine tool, it is necessary to solve some specifi c aspects of material removing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002370_ijvas.2011.041387-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002370_ijvas.2011.041387-Figure1-1.png",
        "caption": "Figure 1 A planar serial manipulator model",
        "texts": [
            " The model is generally considered so as the end-effecter can possess a constant connection with a flat frictional environment. In following, the results are applied to a case study approach to explain the derivation of the equations of motion by this formulation. In this section, the kinematic and dynamic model of an n-DOF planar manipulator with revolute joints is developed, which is included a closed-loop kinematic chain by the connection between the end-effecter (tip of last link) and the environment. The manipulator is supposed to move in the vertical plane x-y, as shown in Figure 1. Let \u0398 be an n-dimensional vector inclusive of joint coordinates \u03b8i, and \u03a6 be an n-dimensional vector with components (i = 1\u2026n): 0 10 and .i i i\u03d5 \u03d5 \u03d5 \u03b8\u2212= = + (1) In fact, the \u03d5i is the angle of ith link with respect to positive direction of horizontal axes. Moreover, the length, mass and moment of inertia about centroid of the ith link are denoted by li, mi and iI , respectively. Let Pp and vp be position and linear velocity of the pth joint, respectively, hence 1 1 (cos sin ) p p i i i i l \u03d5 \u03d5+ = = +\u2211P i j (2) 1 1 ( sin cos ) p p i i i i i l \u03d5 \u03d5 \u03d5+ = = \u2212 +\u2211v i j (3) where, by considering that the base of manipulator has been fixed, it is held: P0 = v0 = 0",
            " ( ) sin sin ( (cos sin ) ) ( ) n n n p p p p p p p p p p p p p n n n p p p p p p p p p p p p p p p W N l l l N N Q \u03b4 \u03b4 \u03b4 \u03c4 \u03b4 \u03d5 \u03d5 \u00b5 \u03d5 \u03b4\u03d5 \u03d5 \u03b4\u03d5 \u03c4 \u03c4 \u03d5 \u00b5 \u03d5 \u03b4\u03d5 \u03c4 \u03c4 \u03bb \u03b4\u03d5 \u03b4\u03d5 + \u2212 = = = + + = = =   = + = \u2212 + +    = \u2212 + + \u2261 \u2212 + \u2261 \u2211 \u2211 \u2211 \u2211 \u2211 \u2211 \u03c4 \u03b8 F P (19) where \u03c60 = \u03b4\u03c60 = 0 and \u03c4n+1 = 0. Moreover, \u03bbp is defined as: (cos sin ).p p p pl\u03bb \u03d5 \u00b5 \u03d5= + (20) It should be noted that the \u03a6 is not an independent generalised coordinates set and they are correlated to each other by geometric constraint, i.e., the connection between end-effecter and ground that causes the loop of kinematic chain is closed. In fact, N may be considered as Lagrange\u2019s multiplier appeared due to the geometric constraint. As shown in Figure 1, it is clear that: 1 sin . n p p p l d\u03d5 = =\u2211 (21) Now, using the Lagrange formulation, equations of motion of the dynamic system can be written as: d . d k k k k T T V Q t \u03d5 \u03d5 \u03d5  \u2202 \u2202 \u2202\u2212 + = \u2202 \u2202 \u2202  (22) By substituting equations (14), (17) and (19) into equation (22), we will have: 1 1 1 1 1 1 1 1 12 ( , ) ( , ) ( , ) cos 2 2 2 ( ), 1, , . pn k n k k k p k p k k p k i j p k k k k m I m M i k M p k m M j k m gl N k n \u03d5 \u03d5 \u03c4 \u03c4 \u03bb \u2212 \u2212 = + = = = + +    + + + + +       = \u2212 + = \u2211 \u2211 \u2211 \u2211 \u2026 (23) In this stage, the derivation of the equations of motion for an n-DOF planar manipulator is completed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002989_2425296.2425316-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002989_2425296.2425316-Figure7-1.png",
        "caption": "Figure 7: Projection of the repulsive force generated on each spherical element on the manipulator arm to an equivalent force at the wrist/elbow.",
        "texts": [
            " When two spheres are in contact each other, a virtual repulsive force for collision avoidance is generated between them. Referring to Figure 6, the direction of the virtual force is directed along the line passing through the centres of the two spheres and the magnitude of the force F is shown below: Fij = K|dij \u2212 (ri + rj)| (6) where K is a spring constant. During collision of robot segments, consider the case of a repulsive force Fr1 being generated on the upper arm at a distance of lr1 from the shoulder joint as shown in Figure 7. The force Fr1 can be transformed into an equivalent force at the elbow using the expression below: Felb = Lr1 L1 Fr1 (7) Similarly, a repulsive force Fr2 generated on the forearm at a distance of lr2 from the elbow joint can be transformed into an equivalent force at the wrist by: Fwr = Lr2 L2 Fr2 (8) For multiples collisions between bounding spheres, Felb and Fwr become a summation of the equivalent forces at the elbow and wrist respectively. As shown in Figure 8, the resultant force Felb produces a moment about the centre of the circle which traces all the possible elbow positions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001622_college.math.j.42.4.289-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001622_college.math.j.42.4.289-Figure8-1.png",
        "caption": "Figure 8. Angles in the cone.",
        "texts": [
            " Now L is the straight line through A parallel to BD, hence L also passes through the center C of the conic, the midpoint of OD, from the midpoint theorem applied to triangle OBD. We know that |CS| = |CS\u2032| since the foci of a conic are equidistant from the center. We would like to show that |CS| = |CA|, which would mean that S and S\u2032 were on the strophoid of L with pole O and fixed point A. From a standard result (in [1, p 226], for example), the eccentricity e of a conic section is given by e = sec\u03b1 cos t , where \u03b1 is the semi-vertical angle of the cone and t is the angle between the plane section and the axis of the cone (see Figure 8). Also, the focal length |CS| = ae, where a = |CO| is the length of the semi-major axis. Hence |CS| = a sec\u03b1 cos t . Let PQ be the diameter of the circular section of the cone through C and let OF be the line through O parallel to the axis of the cone, meeting PQ at F (see Figure 9). Now ABPC is a parallelogram, so that |CA| = |PB|. But |PB| = |QO| = |FO| sec\u03b1 = a cos t sec\u03b1. Therefore |CA| = |CS|, as required. The proof for a hyperbolic section is left as an exercise. In the parabolic case, the major axis OD is parallel to L , so that the above construction fails"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.20-1.png",
        "caption": "Figure 10.20 Stator and rotor current in \u03b1, \u03b2 coordinates",
        "texts": [
            "47 becomes dimr\u03b1 dt = 1 Tr (is\u03b1 \u2212 imr\u03b1) \u2212 imr\u03b2 \u00b7 \u03c9 dimr\u03b2 dt = 1 Tr (is\u03b2 \u2212 imr\u03b2) + imr\u03b1 \u00b7 \u03c9 (10.49) Stator current can be easily transferred from the abc system to \u03b1\u03b2 system. Equation 10.49 can be implemented discretely in the time domain, therefore, imr\u03b1 and imr\u03b2 can be observed. Once this has been done, imr and \u03b4r can finally be calculated: imr = \u221a i2 mr\u03b1 + i2 mr\u03b2, cos(\u03b4r ) = imr\u03b1/imr , sin(\u03b4r ) = imr\u03b2/imr (10.50) where \u03b4r is the angle between the fictitious current imr and the stator current iS\u03b1 as shown in Figure 10.20. If the frame is chosen such that B is aligned with \u03bbR as shown in Figure 10.21, imr will only have real components. Therefore this rotor equation can then be decomposed into its direct and quadrature components as follows: isd = iS\u03b1 cos \u03b4r + is\u03b2 sin \u03b4r isq = \u2212iS\u03b1 sin \u03b4r + is\u03b2 cos \u03b4r (10.51) From Equation 10.45, when is is decomposed to d\u2013q components, the equation can be written as follows: imr \u2212 isd + Tr \u00b7 pimr = 0 (10.52) \u2212isq + Trimr \u00b7 p(\u03b4 \u2212 \u03b8) = 0 (10.53) From Equation 10.52 it can be seen that imr is only related to isd "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure1-1.png",
        "caption": "Figure 1: Procedure for Hohlraum assembly",
        "texts": [
            "0 of NIF (National Ignition Facility) targets [5, 6], the axiality and axial angle of subassemblies for Hohlraum assembly are less than 2 \u03bcm and 0.5\u00b0, which are hard to achieve. The precision robotic assembly system developed for Hohlraum assembly is necessary [7]. Hohlraum assembly is one important procedure of fusion ignition target assemblies, including Hohlraum inserted into TMP, LEH (Laser Entrance Hole) inserted, and Tent and Tamping gas tube assembled [6]. The procedure of Hohlraum assembly is shown in Figure 1. In the procedure of Hohlraum assembly, Hohlraum insertion is the hardest. Sticking Hohlraum and TMP will influence the DT (deuterium and tritium) ice layer homogenized. The best way to fix them is shrink assembly. Owning to the coefficient of thermal expansion of Al (aluminum) is larger than Au (aurum), Hohlraum and TMP will be pressed more after freezing to 20 K. The shrink range of Hohlraum and TMP determined is very important. Based on FEM (finite element method), the max displacements of pressing and freezing are analyzed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure18-1.png",
        "caption": "Figure 18. Normalised Stress at RSRB region (Linear analysis)",
        "texts": [
            " In linear analysis, high stress is observed in the fillet region of External flitch due to load transfer only through bolts and non-consideration of contact between External flitch top flange and FSM top flange. This is not so in case of local non-linear analysis as contact is defined. Figure 16 and 17 shows comparative stress plot at RSMB region between linear and local non-linear analysis respectively. Similar trend is observed in the fillet regions and high stress is shifted to other locations due to consideration of contact. In linear analysis, high stress is concentrated around bolt holes as the load transfer happens through bolts only. Figure 18 and 19 shows comparative stress plot at RSRB region between linear and local non-linear analysis respectively. Similar trend is observed in the fillet regions and high stress is shifted to other locations due to consideration of contact. In linear analysis, high stress is concentrated around bolt holes as the load transfer happens through bolts only. It is also observed that stress levels on bottom portion of FSM close to RSRB bracket remains same in both the analysis. This might be due to contact does not exist in that region"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure1-1.png",
        "caption": "Figure 1. FE model of truck chassis assembly.",
        "texts": [
            " Welds are simulated using rigid elements. Thin shell elements are used to model FSM, cross members and flitches. Spring mounting brackets are meshed using linear tetrahedral elements. Aggregate masses, load body and payload masses are modeled as lumped mass at respective CG locations and are connected to the structure using RBE2 and RBE3 elements. The pin joints in the suspension are simulated using CBAR and RBAR elements to allow rotational degrees of freedom. FE model of frame assembly is shown in Figure 1. The wheel points are constrained such that the structure is subjected to minimum constraints. Linear static analysis is carried out to evaluate the stresses on FSM for vertical 3g loading condition. FE model is solved using MSC / Nastran and results are post-processed using Altair / Hyper view. Figure 2 shows the stresses on the chassis assembly. The below mentioned regions are identified as critical regions where linear assumptions are invalid because load transfer happens through surface contact interactions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002935_kem.552.248-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002935_kem.552.248-Figure3-1.png",
        "caption": "Fig. 3 Simulation model of the FTS",
        "texts": [
            " Natural frequency of one spring system could be: ) 3 /( 0 0 m Mk +=\u03c9 (1) Among them k0 is the elastic coeffiency of the flexible hinge; M is the quality of the object; m0 is the quality of the spring. Increase a parallel spring, the frequeny could be: ) 3 /()( 10 10 mm Mkk + ++=\u03c9 (2) Among them, k1 is elastic coeffiency of spring that can be changed; m1 is the quality of the spring. From equation (2), it can be seen that if k1 changed the frequency of the FTS is also changed. The simulation model can be seen in Fig3. Material is Stainless steel No.304 that data is shown in Table 1: Results of top six order model analysis could be seen in Table 2: Fig.2 Simplified FTS model F K1/2 Ko2 Ko3 Ko4 Ko1 K1/2 (a) (b) F Ko K1 with spring that stiffniss is 300, 1000 and 3000N/mm; initial compression is 3mm. From simulation it can be seen that : 1 frequency is changed when intall a parallel spring. 2 the more the stiffniss of the spring, the higher the frequency will be. The FTS structure includs frame rack and flexible hinges which is cut by a piece of metal"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002280_synasc.2013.9-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002280_synasc.2013.9-Figure1-1.png",
        "caption": "Figure 1. Regular pentagon EHGIF obtained by knot fold",
        "texts": [
            " In this paper we are interested in the concrete shapes created by knotting the paper tape. Hence, we are less concerned with the topological and combinatorial aspects of knots, but are more with the methods of the construction. Furthermore, since a paper tape of a rectangular shape can be constructed from the square sheet of paper (i.e. origami) by repeated folding, we will consider a rectangular origami from the outset. Subsequently, we simply call it a tape or origami. We depict a knotting operation in Fig. 1. In the figure we make one simple knot. When the height of the tape, i.e \u2223AD\u2223(= \u2223BC\u2223), is infinitesimal and both ends of the tape are connected, the tape can be viewed as a closed curve, i.e. an object of study in the theory of knots. The knot with 3 crossings is the most basic one. It is denoted 31 in the Alexander-Briggs notation. When the height is finite, each crossing or the collection of the crossings have a certain polygonal shape if projected onto the plane. It is called a polygonal knot",
            " The reason to have \u201cCase \u2192 1\u201d in this case is that we have two solutions. In general there are two fold lines in (O5) fold. To satisfy the conditions of the regularity, we have to pull the both ends AB and CD of the tape outwards in Fig 5, so that the knot becomes tight. This fastening operation is none of Huzita\u2019s basic operations. A question arises naturally whether we can make the knot using the Huzita\u2019s set of basic operations. This is one of the topics of the next section. Let us look at the initial origami Fig 1(a) again. We may fix the point E on the edge AB. The location of the point F on the edge CD relative to the location of the point E completely determines the subsequent shapes if we obey the same method of the folds discussed in the previous section. Suppose that we slide the point F along the edge CD towards D of the tape in Fig. 3. This small move propagates in the subsequent three steps of the folds and brings the points K and I closer. The fastening of the knot amounts to those four operations i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure8-1.png",
        "caption": "Fig. 8 The actual device",
        "texts": [],
        "surrounding_texts": [
            "Some discussions of the new design were done during the exercise. The shape of the locking pin was cylindrical, and the locking top matched with a groove in the holding device, which was also in the cylindrical shape, thus the contact surface was just a short line in the vertical direction and a small circular area, and this area was responsible for pressing the holding device, this made lots of attrition, the attrition would shorten the service life of the locking pin. If the situation went on, the quality of the connection will be seriously impacted. Every parts of the connecter were made from the same material, as time went by, the friction force on the contact surface of the two connected parts would be increased, and it might lead to make the connector unworkable. Besides, the weight of the whole structure needed to be reduced by changing shape of the holding device and the locking structure. Further improvement of the design would be to perfect the parts mentioned above. \u2164. CONCLUTION In this paper the typical kinds of mechanical quick connector used in medical devices were enumerated and analyzed. From those examples we can conclude some clues for designing a mechanical quick connector for medical equipment. One clue is that using traditional screw joint connection, if the connection doesn\u2019t need to connect very quickly; another clue is that taking examples from the ideas of other mechanical connectors used in other fields, such as pneumatic joints and electronic connectors etc. This kind of designs usually used in the situation that have a strict demand for the short connect-time. Then a new kind of connector was designed. The machining was done according to the design. The result showed that the design was successful."
        ]
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure3-1.png",
        "caption": "Fig. 3. PAMD Extended View",
        "texts": [
            " The pros or +1\u2019s of the Design 3 are the cost of components and ease of manufacturing due to it being similar to a foldable bike. Its concept and weighted concept rating is -1 and 0, respectively. Hence, Design 1 was selected as the final concept design to refine and perform the engineering analysis. The detailed CAD designs of the PAMD using the NX 7.5 software are shown in Figures 1 and 2 with dimensions. The final design specifications are shown in Table 3. The final design of the PAMD is shown in Figure 3. III. ENGINEERING ANALYSIS Theoretical stress calculations were completed on the front axle of hub and the motor shaft of the electric bike kit. The components are loaded with maximum 1000 lbf which includes a safety factor of 5. The results of the shear and bending moment diagrams for the hub/axle and the motor shaft showed maximum bending stress of 48,414 and 53 psi, respectively. The maximum shear stress for each is 24,220 and 3,872 psi, respectively. Stainless steel or aluminum alloy at yield strength of 75,000 psi and 60,000 psi, respectively, would be a great choice of material for these components [8]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003081_gt2012-69967-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003081_gt2012-69967-Figure9-1.png",
        "caption": "Figure 9. Close-up of shaft face showing maximum angular displacement at the periphery of the base diameter",
        "texts": [
            " Geometric elements selected as parameters that can be modified would consist of the base shaft diameter, web thickness and height, and the number of webs. In addition to the torsional stiffness calculation, a solid model also allows for the verification of the motor core weight and the inertia. Fig. 7 provides the meshed FEA model used to determine the rotational deflection used in the stiffness calculation. Figure 8 shows the torsional deflection of the shaft when it is subject to the face constraint at one end and the torque at the other. Figure 9 shows a close-up of the face undergoing maximum deformation. Four points are selected along the periphery of the diameter to obtain an average rotational deformation. The angular deformation is then determined from Eq. (14). l r   (14) It was observed that the torsional stiffness is generally higher as the number of spider bars as well as individual spider bar thickness increases. These stiffness values will later be compared with the other methods described in this paper. An additional benefit of generating a solid model and subsequent FEA analysis of a webbed shaft is the estimation of stress concentration factors (SCFs) that are commonly used in torsional analyses"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002326_tmag.2013.2247984-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002326_tmag.2013.2247984-Figure3-1.png",
        "caption": "Fig. 3. Two special cases for Type-3 response.",
        "texts": [
            " In fact, this is the assumption made in [1]. Type 3: . The type-3 response is between the two extremes of type-1 and type-2 responses. It is difficult to characterize rotor response in situ for this regime, because many scenarios may occur. Two special cases are discussed as follows. Special Case A: This exists when the frequency of interest is near the half-speed whirl satisfying and the proof mass is chosen so that . In this case, the denominator in (8) nearly vanishes, implying a very small response at the lower bearing; see Fig. 3. For this case, it will be very difficult to extract bearing coefficients from (10) and (11) because the left sides of the equations are similar to a zero-divided-by-zero case. Special Case B: This exists when the frequency of interest and the proof mass are chosen such that , the rocking of the spindle could become negligible; see Fig. 3. For this case, the related terms in (10) and (11) may be ignored. Understanding of the three types of physical response can be very helpful when extracting the bearing coefficients. For example, it can help us to design the size of the proof mass so that and are all significantly measurable in experiments. It can also serve as a guideline to evaluate whether or not experimental results make sense. Now let us assume that , and have been measured experimentally. Then, the left side of (10) or (11), defined as and , respectively, can be calculated and plotted as a function of (in magnitude and phase)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002441_amr.544.286-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002441_amr.544.286-Figure9-1.png",
        "caption": "Fig. 9 Typical Mises stress cloud map for partials of the outer ring, rollers, and inner ring",
        "texts": [
            " And then, in stage II the local part around the bearing including additionally pieces of axle and vice-frame is calculated in the manner of taking the stress-stain parameters on the cut sections as constraint conditions and the bearing three parts, i.e. outer ring, rollers and inner ring, as contacting physically. Results and Discussions. A piece of dynamic balance force history for the merged part were first deduced from on-line inspected wheel-rail contact force spectra on a straight railway line with car velocity of 120 km/h. Typical Mises stress cloud map for partials of the outer ring, rollers, and inner ring is given in Fig. 9. Maximum equivalent RCF stresses of the outer ring, rollers, and inner ring are given in Table 3. Combining the analyzed results for same bearing on a common type of freight car [6], it is cleared that \u27a3 Equivalent stresses at the location nearby seal seat groove of outer ring are 1.335 and 1.488 times of those at the location nearby inner side of inner ring for the present car and the common car, respectively. It indicates that the design of out ring may mismatch inner ring for the bearing. The bearing may normally appear an earlier failure from the location nearby seal seat groove of outer ring"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure2-1.png",
        "caption": "Figure 2: Displacements concentration based on FEM",
        "texts": [
            " Tel: 86-816-2494151; 6th International Symposium on Advanced Optical Manufacturing and Testing Technologies: Design, Manufacturing, and Testing of Smart Structures, Micro- and Nano-Optical Devices, and Systems, edited by T. Ye, X. Luo, S. Hu, X. Bao, Y. Li, Proc. of SPIE Vol. 8418, 841819 \u00b7 \u00a9 2012 SPIE \u00b7 CCC code: 0277-786/12/$18 \u00b7 doi: 10.1117/12.977645 Proc. of SPIE Vol. 8418 841819-1 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 04/21/2013 Terms of Use: http://spiedl.org/terms of Hohlraum and TMP is 0-4.9 \u03bcm. The displacement concentration of Hohlraum and the aluminum tube of TMP by FEM are shown in Figure 2. There are three components in the precision robotic assembly system for Hohlraum assembly, namely TMP microoperation stage, Hohlraum micro-operation stage and detection system online. The composition and layout of the system are shown in Figure 3 and 4. Proc. of SPIE Vol. 8418 841819-2 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 04/21/2013 Terms of Use: http://spiedl.org/terms TMP manipulation stage is used for holding, moving and adjusting TMP. The stage is comprised of XYZ liner stages and TMP holder (Figure 5)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002547_amm.394.427-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002547_amm.394.427-Figure1-1.png",
        "caption": "Figure 1. Schema with forces and moments in the airplane.",
        "texts": [
            " The longitudinal model is obtained from the full model, assuming the roll and yaw movements are zero, as shown in [7,8], and is given by four differential equations expressed in the body frame and wind as follows: cos sin sin cos /!\" where the D and L represent the aerodynamic forces of drag and lift respectively, the term is the relative velocity vector in the wind direction, the thrust force vector is represented by F$, the pitch angle is represented by \u03b8, the angle of attack is denoted by \u03b1, the moment of inertia about the y axis is defined by J& and the pitch moment is denoted by . The schema of the longitudinal airplane model with forces and moments, where ' , Represents the angle of trajectory is shown in figure 1. If the equation (3) is derived with respect to time, and replaced in equation (4) yields the following expression for the angular acceleration: ( /!\" (5) The Pitch moment of the airplane can be expressed as the sum of moments generated by the wing ( )), the horizontal stabilizer ( *) and the damping moment ( + ) due to the variation of pitch as shown follows: (1) (2) (3) (4) ) * + ,-. /0 \u0305/2 (6) where the dimensionless coefficient of the pitch moment is given by: ,- ,-34 ,-5 ,-6789 ,-: ;\u0305 <= (7) where ,-34 is the pitch moment coefficient at the mean aerodynamic center of the wing, ,-5 is the pitch moment coefficient of the wind due to the change in alpha (angle attack)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003881_etfa.2011.6059220-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003881_etfa.2011.6059220-Figure3-1.png",
        "caption": "Figure 3. Trajectory computed by motion comp",
        "texts": [
            " Algorithm 4 Heuristic for degenerate movement if s \u00b7 s\u03c0 < 0 then v\u2217p \u2190 \u03bbv\u2217p {0 < \u03bb < 1} if v\u2217p < v\u2217p,min then {Could not find a viable value for v\u2217p} return with fail code end if v (i) p \u2190 s\u0302 \u00b7 v\u2217p Increment i continue {Restart loop with new value of v\u2217p} end if Algorithm 1 computes the parameters that define the movement, namely p\u03b1, p\u03c9 and vp. The actual computation of the points in the trajectory is performed by the algorithm motion comp, described in algorithm 5. At the end of motion comp, pp is a vector that contains all points in the trajectory. Figure 2 presents the major elements that define the movement: \u2022 pi, vi, pf and vf , that were input to mov2d; \u2022 p\u03b1, p\u03c9 and vp, that were computed by mov2d. Figure 3 corresponds to the same trajectory, after superimposing the points computed by motion comp. The result is a set of points that the robot\u2019s control system can track an follow, as shown in figure 4. We present here an heuristic to compute the points in a trajectory in a 2D space based on the trapezoidal velocity profile, where the movement occurs in three phases: Algorithm 5 motion comp: compute points in a trajectory {Elements for phase 1} \u0394v1 \u2190 vp \u2212 vi {\u0394v in phase 1} t1 \u2190 |\u0394v1| a\u2217 1 {Time to complete phase 1} n1 \u2190 t1 Ta {Number of cycles in phase 1} a1 \u2190 \u0394\u0302v1a \u2217 1 {Elements for phase 3} \u0394v3 \u2190 vf \u2212 vp {\u0394v in phase 3} t3 \u2190 |\u0394v3| a\u2217 3 {Time to complete phase 3} n3 \u2190 t3 Ta {Number of cycles in phase 3} a3 \u2190 \u0394\u0302v3a \u2217 3 {Start computing the points} p \u2190 pi; v \u2190 vi; {Initial values} {Phase 1} for n = 1 to n1 do v \u2190 v + a1Ta p \u2190 p + vTa pp \u21bc p {A \u21bc b: append b toA} end for {Phase 2} p \u2190 p\u03b1; v \u2190 vp {Initial point for phase 2} pp \u21bc p for n = 1 to n2 do p \u2190 p + vTa pp \u21bc p end for {Phase 3} p \u2190 p\u03c9 {Initial point for phase 3} pp \u21bc p for n = 1 to n3 do v \u2190 v + a3Ta p \u2190 p + vTa pp \u21bc p end for p \u2190 pf {Last point in the movement} pp \u21bc p initial acceleration from initial velocity to a plateau speed (phase 1), travelling in uniform motion at plateau velocity (phase 2) and acceleration from plateau velocity to final velocity (phase 3)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.69-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.69-1.png",
        "caption": "Figure 2.69. Squirrel-cage rotor induction machine: (a) squirrel cage of rotor; (b) squirrel cage",
        "texts": [
            " Their terminals are connected to a passive external circuit through slip rings, in case of wound rotor, or are short-circuited internally when the rotor has a simple or double squirrel cage type construction. Induction motor rotors are made in two forms: short-circuited (squirrel cage) and with slip rings. The first of these is simpler in construction and is more frequently used. The winding of such a rotor is in form of a cylindrical cage (the so-called \u201csquirrel cage\u201d) of copper or aluminum bars, short-circuited around the ends by two rings. The bars of this winding are placed without insulation in the slots of the rotor (Figure 2.69a). Another method in use is to pour molten aluminum into the slots of the rotor. A slip-ring rotor, also known as a phase-wound rotor, has a winding made of insulated wire (Figure 2.70a) and in most cases in three-phase wye connected. The free ends of this winding are brought out to the slip rings on the rotor shaft. Brushes bear on the slip rings and connect the rotor winding to a three-phase rheostat (Figure 2.70b). This system permits the resistance of the rotor circuits to be varied. This is very important in starting the motor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002414_amm.99-100.857-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002414_amm.99-100.857-Figure1-1.png",
        "caption": "Figure 1 Illustration of tapered gap",
        "texts": [
            " Force direction and magnitude of 2D bearing On the basis of Sommerfeld hypothesis and the theory of the immensity bearing, the Reynolds equation could be change to differential equation as follows (one-dimensional Reynolds equation): The direction of the resultant force of oil-film F (equal to \u03b7F ) is vertical to the direction of eccentric ratio e at all time. Though the value of e is different, the angle between the direction of the resultant force of oil-film and axis center line is always 90 0 . As shown in figure 1: for tapered gap, assumed that the pressure along the direction of the thickness of oil-film is constant, ignore the convection item, assumed only 2 2 y u \u2202 \u2202 \u03b7 and 2 2 y v \u2202 \u2202 \u03b7 are correlative with viscosity, than the N-S equation can be rewritten as follows: ( )       \u2212+ \u2202 \u2202       + + \u2202 \u2202       + =      \u2202 \u2202 \u2202 \u2202 +      \u2202 \u2202 \u2202 \u2202 12 2121 33 22 12 VV z hWW x hUU z ph zx ph x \u03b7\u03b7 (3) This equation is the Reynolds equation for the tapered gap. The 2D model which could be solve for the tapered gap is shown as figure 2 shows and dimensions are as follow: B=0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001687_978-3-7091-1187-1_5-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001687_978-3-7091-1187-1_5-Figure1-1.png",
        "caption": "Figure 1. Coordinate axis, reference displacements and beam dimensions.",
        "texts": [
            " In what concerns plasticity, the book of Kojic\u0301 and Bathe (2005) provides a description of the concepts here employed. Nevertheless, readers interested in inelastic problems may wish to additionally consult other books, including (Owen and Hinton, 1980) and (Simo and Hughes, 1998). Displacements are here defined with respect to a stationary reference frame, i.e., in a \u201ctotal Lagrangian\u201d approach. The displacement field is based upon the following assumptions: - the beam vibrates only in one plane (we designate this plane as x1x3, represented in Figure 1); - the beam is initially straight; - the beam cross sections remain plane; - cross sections are free to rotate about an axis perpendicular to the plane where motion takes place; - the beam thickness is moderately small in comparison with the length, agreing with the definition of a beam as an elemental structure, but it is not necessarily very small, because rotations of the transverse section are considered. The assumptions above lead to a first order shear deformation - also known as Timoshenko - model and the displacement field is given by u1 (x1, x3, t) = u01 (x1, t) + x3\u03b8 0 (x1, t) (1) u3 (x1, x3, t) = u03 (x1, t) (2) where ui(x1, x3, t) represents the displacement component along axis xi and \u03b80 (x1, t) represents the cross section rotation about x2. Superscript 0 indicates the longitudinal axis, x1, which passes through the cross sections centroids, when the beam is straight. The three reference axes and beam dimensions (length , width b and thickness h) are represented in Figure 1. In addition, it is assumed that the displacements are moderately large and Green strain tensor (Chia (1980), Fung and Tong (2001)) \u03b5ij = 1 2 ( \u2202ui \u2202xj + \u2202uj \u2202xi + \u2202uk \u2202xi \u2202uk \u2202xj ) , i, j, k = 1, 2, 3 (3) is employed. In the case at hand, the most important nonlinear term is( \u2202u3 \u2202x1 )2 and only this will be considered (a von Ka\u0301rma\u0301n approach, Chia (1980)). Hence, the longitudinal strain and the transverse shear engineering strain, \u03b313 (x1, t), which is twice the tensorial shear, are the following \u03b511 (x1, x3, t) = \u2202u01 (x1, t) \u2202x1 + 1 2 ( \u2202u03 (x1, t) \u2202x1 )2 + x3 \u2202\u03b80 (x1, t) \u2202x1 (4) \u03b313 (x1, t) = 2\u03b513 (x1, t) = \u2202u03 (x1, t) \u2202x1 + \u03b80 (x1, t) (5) The longitudinal strain can be written in the following form, which is advantageous to define the stiffness matrices (Ribeiro (2001)): \u03b511 (x1, x3, t) = \u230a 1 x3 \u230b({ \u03b5 L (x1, t) \u03b5b L (x1, t) } + { \u03b5 NL (x1, t) 0 }) (6) In equation (6) three strain components appear: the linear longitudinal strain, \u03b5 L (x1, t), the bending strain, \u03b5 b L (x1, t), and the geometrically non- linear longitudinal strain, \u03b5 NL (x1, t)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.72-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.72-1.png",
        "caption": "Fig. 2.72 4PaPaC-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\Pa\\\\C",
        "texts": [
            "22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No. 58 T ab le 2. 4 L im b to po lo gy an d co nn ec ti ng co nd it io ns of th e fu ll y- pa ra ll el so lu ti on s w it h no id le m ob il it ie s pr es en te d in F ig s. 2. 73 , 2. 74 , 2. 75 , 2. 76 , 2. 77 , 2. 78 ,2 .7 9, 2. 80 ,2 .8 1, 2. 82 ,2 .8 3, 2. 84 ,2 .8 5, 2. 86 ,2 .8 7, 2. 88 ,2 .8 9, 2. 90 ,2 .9 1, 2. 92 ,2 .9 3, 2. 94 ,2 .9 5, 2. 96 ,2 .9 7, 2. 98 ,2 .9 9, 2. 10 0, 2. 10 1, 2. 10 2, 2. 10 3, 2. 10 4, 2. 10 5, 2. 10 6, 2. 10 7, 2. 10 8, 2. 10 9, 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.35-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.35-1.png",
        "caption": "Fig. 2.35 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PaRPP (a) and 4PaPRP (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology Pa\\R||P\\P (a) and Pa||P\\R||P (b)",
        "texts": [
            "21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.39-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.39-1.png",
        "caption": "Figure 12.39 Path of the top symmetry axis on a sphere when \u03d5\u0307 (t) is changing between negative and positive values.",
        "texts": [
            " The superposition of periodic changes of \u03b8(t) onto a precession about the Z-axis with a periodically changing angular velocity \u03d5\u0307 (t) generates a wavy motion of the z-axis. It can be visualized on a sphere with the center at the fixed point X Z \u03c8 \u03d5 Y \u03b8 z The general paths of motion of the symmetry axis of the top are illustrated in Figures 12.37\u201312.39. Figure 12.37 depicts the situation in which \u03d5\u0307 (t) is changing between two positive extreme values. Figure 12.38 shows the path of the z-axis when \u03d5\u0307 (t) is changing between negative and positive values. Figure 12.39 depicts the situation in which \u03d5\u0307 (t) is changing between zero and a positive maximum. The periodic dance motion of the top with \u03b8 (t) which is superimposed on the precessional motion is called nutation. Example 756 Special Cases of Top Dynamics Some special cases of top dynamics have simpler equations of motion with simpler interpretations. Consider the top kinematics shown in Figure 12.35 and its dynamic equations of motion (12.603)\u2013(12.605): 1. \u03c93 = 0. This case is equivalent to a planar pendulum in which \u03b8 is the only time-dependent variable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003464_amm.397-400.176-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003464_amm.397-400.176-Figure4-1.png",
        "caption": "Fig. 4 Planet carrier with shaft collar and main bearings in position",
        "texts": [
            " The planet carrier transmitted torque from the spur gear to the planet gears and each planet gear was cantilevered off bearing pins. It was therefore essential that the carrier assembly was not only strong enough to support the applied loads but also stiff enough to allow only minimal deflections under load. Large deflections would cause misalignment between the planet gears and the ring gear causing increased vibration levels and tooth wear. The planet carrier was therefore analysed using Finite Element Analysis to ensure the required stiffness was achieved. The planet carrier was manufactured from 6061 Aluminium Alloy. Fig. 4 shows the two main bearings which locate the planet carrier on the shaft and allow it to rotate independently. A shaft collar was used to mount the bearings on the driven shaft and locate the planet carrier axially. The collar also aligned the planetary stage with the driven shaft. Any misalignment in the collar or the main bearings would be amplified due to the diameter of the carrier and would cause a significant level of precession in the other components on the planetary stage. The main carrier bearings fit over the collar and were held in place with two circlips. The collar assembly is shown in Fig. 4. The large spur gear was bolted to the carrier using three 8mm shoulder bolts. Shoulder bolts were used to ensure accurate location of the spur gear on to the carrier as any misalignment would affect its meshing with the pinion gear on the driving shaft. Three steel pins were press fit into recesses in the carrier. One of the two planet bearings on each gear fits over this pin and inside the machined hole in the planet gear as shown in Fig. 5. The other planet bearings fits into the opposite side of the planet gear with a steel bush located between the two bearings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003486_icicta.2011.588-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003486_icicta.2011.588-Figure1-1.png",
        "caption": "Figure 1 The structure of Planetary gear train If gear 3 relatively fixed, the meshing",
        "texts": [
            " M HP Said planetary gear transmission of friction loss, and the value is, the bigger the instructions of planetary gear transmission is lower. In order to show planetary gear transmission power and mesh power, the relationship between the power factor into meshing[3].. \uff1d H S S P P Positive and negative number mechanism of planetary gear transmission, refers to the transformation agency respectively for the transmission ratio 1 H ni positive number or negative number mechanism gear train. In the following figure 1 the turnover gear train. 1, 3 as the center of gear wheel, component H for planet shelf (turning arm). power factor: \uff1d H S S P P \uff1d 1 1 HP P \uff1d  1 1 1 1 HT n n T n  \uff1d1\uff0d 1Hi ............\uff084\uff09 Type: 1T - gear 1 input torque; 1n \uff0dgear 1 input speed ; Hn - planet shelf input speed. By turnover gear train transmission ratio formula knowable:     1 13 3 H H H n n i n n    , because gear 3 fixed, so [2] :     1 13 3 H H H n n i n n    \uff1d     1 11 0 H H H n n i n     , namely 1Hi \uff1d1\uff0d 13 Hi \uff0cWill it into formula (4)\uff0c available \uff1a \uff1d 13 13 1 H H i i  \uff1d 1 1 1H H i i  "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.3-1.png",
        "caption": "Figure 3.3 The external forces resolved in the wind axes frame, (iv, jv, kv), with flight-path angle, \u03c6, and bank angle, \u03c3",
        "texts": [
            " The condition in which the flight direction is confined to the plane of symmetry also leads to a zero sideforce and is called coordinated flight. To achieve coordinated flight and hence the maximum aerodynamic efficiency, an aircraft requires precise attitude control \u2013 a topic to be discussed later. In the absence of a sideforce, the only way the aircraft can be made to turn horizontally in a coordinated flight is by banking the wings, that is, tilting the plane of symmetry such that it makes an angle, \u03c3, (called the bank angle) with the local vertical plane, (iv, kv), as shown in Figure 3.3. In such a case, the lift force, L, acts in the (jv, kv) plane, thereby creating a centripetal force component, L sin \u03c3, along jv (Figure 3.3). In order to achieve the maximum fuel economy, the direction of thrust, fT, must lie in the plane of symmetry, and should be nearly aligned with the flight direction, iv. Thus, we assume that the thrust makes an angle, \u03b5 (called the thrust angle) with iv, as shown in Figure 3.3. Except for a few special aircraft that are capable of thrust vectoring (i.e., changing the thrust angle in flight by thrust deflection), all conventional aircraft have a small but fixed thrust angle due to the engines being rigidly bolted to the airframe. In such a case, the thrust, fT, approximately acts along the longitudinal axis of the aircraft. The sum of aerodynamic, fv, and thrust, fT, force vectors for coordinated flight resolved in the wind axes can be expressed as follows: Fv = fT + fv = (fT cos \u03b5 \u2212 D) iv + L sin \u03c3jv \u2212 (fT sin \u03b5 + L cos \u03c3) kv, (3",
            " For a rigid aircraft, the aerodynamic and propulsive forces, (X, Y, Z), and moments, (L, M, N), are functions of the relative linear velocity vector, that is, the airspeed, v, as well as the relative flow angles \u2013 the angle of attack, \u03b1, and the sideslip angle, \u03b2, arising out of the orientation of the aircraft\u2019s body axes relative to the velocity vector (or in other words, the linear velocity vector resolved in the body axes).3 We shall employ the relative flow angles, \u03b1, \u03b2, and the geometric bank angle, \u03c3, as Euler angles in order to derive a coordinate transformation between the wind axes, (iv, jv, kv), and the body axes, (i, j, k), as shown in Figure 3.21. Recall from Figure 3.3 that the plane (iv, kv) is the vertical plane, and that (i, k) is the plane of symmetry (the plane containing the lift vector). Therefore, the bank angle, \u03c3, is merely the angle between the two planes (iv, kv) and (i, k) or between the axes jv and j (Figure 3.21). Clearly, the Euler sequence for the necessary coordinate transformation is (\u03b2)3, (\u2212\u03b1)2, (\u03c3)1 (Figure 3.21), resulting in \u23a7\u23a8 \u23a9 iv jv kv \u23ab\u23ac \u23ad = Cv \u23a7\u23a8 \u23a9 i j k \u23ab\u23ac \u23ad , (3.129) where Cv is the following rotation matrix: 3 The vector (X, Y, Z, L, M, N) also depends upon the relative angular velocity vector, (P, Q, R)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure6-1.png",
        "caption": "Fig. 6. Frame Reaction Force Analysis",
        "texts": [],
        "surrounding_texts": [
            "The detailed CAD designs of the PAMD using the NX 7.5 software are shown in Figures 1 and 2 with dimensions."
        ]
    },
    {
        "image_filename": "designv11_100_0001303_pamm.200610036-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001303_pamm.200610036-Figure1-1.png",
        "caption": "Fig. 1 Regularization with a spring, force direction jump",
        "texts": [
            " One can adjust the resolution of frictional forces acting in the contact area by choosing a reasonable set of contact points. Then a friction law for each contact point is formulated. Coulomb friction does not predict the friction force during sticking for multiple contacts. Therefore a regularization of friction similar to the methods in [3] has to be used. However, a standard regularization with springs and dampers fails for multiple contact interfaces and two-dimensional motions. Consider a contact point in sticking state, regularized with a spring (cf. Figure 1). The spring displacement corresponds to a maximal transmittable frictional force \u00b5st. Now a small perturbation with some arbitrary relative velocity vrel is applied. The perturbation causes the contact point to tear off, i.e. the frictional force represented by the spring reaction force (sticking) exceeds \u00b5st and has to be replaced by a force opposing relative velocity (sliding). This is a discontinuous jump of the force direction immediately before (F\u2212 f ) and immediately after (F+ f ) tear-off (cf. Figure 1). A jump of the absolute value of a frictional force is permitted according to some Stribeck-curve. But the discontinuities of friction force orientations may lead to non-accelerating tear-off situations. Furthermore the dependency of a friction force orientation on a small perturbation seems unrealistic from a macroscopic point of view. Here we want to outline the functioning of the proposed regularization method without concentrating on the choice of contact points. As pointed out above, the goal is to use smooth transitions of the frictional force orientations between sticking and sliding states in both directions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003881_etfa.2011.6059220-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003881_etfa.2011.6059220-Figure1-1.png",
        "caption": "Figure 1. CAMBADA\u2019s locomotion system.",
        "texts": [
            " As the only constraints posed on the robot path are those defined by the robot\u2019s dynamics, the proposed heuristics are directed at omnidirectional holonomic robots, i.e., robots that are capable of, amongst others, manoeuvring without affecting the orientation [3]. CAMBADA is the MSL robot soccer team from the University of Aveiro, in Portugal [2]. CAMBADA was MSL World Champion in 2008 and ranked third in the 2009 and 2010 competitions. The robot\u2019s locomotiom system is based on 3 omnidirectional wheels, aligned at 120o (figure 1). These wheels, also called Swedish wheels [1], have rollers that allow for unconstrained movement along the wheel axis (perpendicular to the wheel\u2019s motion). This configuration of the robot wheels allows for holonomic motion. We assume that the robot has some kind of location system that allows it to correctly estimate its position and orientation with respect to a global reference frame with a finite error bound. The algorithm computes and provides the robot a time sequence of spatial positions as points in this global reference frame that the robot will follow"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003082_icems.2013.6754391-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003082_icems.2013.6754391-Figure2-1.png",
        "caption": "Fig. 2. Configuration of a PC-PM motor.",
        "texts": [
            " A motor that can change the magnetic force of a PM was proposed [1]- [12]. In this study, we propose a PM motor that can change the number of magnetic poles by a factor of three, and we present its principle of operation and basic characteristics. Fig. 1 shows the pole changing of motors that are suitable for variable speed operation. When the motor operates with 6 poles, it produces a high torque at low speeds. If the motor changes to 2 poles, it produces a low torque at medium and high speeds. Fig. 2 shows the basic configuration of a polechanging PM (PC-PM) motor that changes the number of magnetic poles by a factor of three. The PC-PM motor has a rotor embedded with PMs that have high coercive force (constant magnetized magnet) and low coercive force (variably magnetized magnet). Moreover, flux barriers are prepared at the end of each magnet to improve the magnetizing characteristics of the PM. Fig. 3 shows the principle of pole changing. The armature winding changes from a 6-pole winding to a 2-pole winding",
            "00 \u00a92013 IEEE (a) 6 poles (b) 2 poles (a) 6 poles (b) 2 poles A finite-element method (FEM) magnetic field analysis was performed using FEM software JMAG to understand pole changing and the characteristics before and after changing. We verified the pole changing of the rotor using magneticfield analysis. Fig. 5 shows the distributions of the magnetic flux density at no load. When the variably magnetized magnets in the poles reverse the polarity of adjacent poles with constant magnetized magnets, the magnetic flux forms a 6-pole distribution, as shown in Fig. 2(a). When the variably magnetized magnets are magnetized in the reverse direction, the magnetic flux forms a 2-pole distribution, as shown in Fig. 2(b). (a) 6 poles (b) 2 poles Fig. 5.Distribution of magnetic flux density at no-load. B. Induced voltage characteristics Fig. 6 shows the induced voltage waveform at no load. The 2-pole waveform represents one cycle and the 6-pole waveform represents three cycles. Pole changing can be confirmed from this result. Fig. 7 shows each harmonic of the induced voltage. The motor changes from 6-pole to 2-pole, and the induced voltage can vary from 100% to 50% because of pole changing. Fig. 8 shows the torque-current phase characteristics"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001340_b978-081551497-8.50005-6-Figure3.41-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001340_b978-081551497-8.50005-6-Figure3.41-1.png",
        "caption": "Figure 3.41 Geometry of the system and degrees of freedom for the displacements of the micropart, with respect to the substrate pad. The shape of the liquid meniscus was computed numerically. Adapted from Greiner et al.[68]",
        "texts": [
            " The surface energies of the water-lubricant interface (46 mJ/m2) and the coating-water interface (52 mJ/m2) show similar values, while the lubricant-coated interface (<1 mJ/m2) is of two orders of magnitude less.[65] This difference of nearly two orders of magnitude in the surface energies is the driving force that assures self-alignment. In our simulations, we calculated the total surface energies for different configurations using the surface evolver software by Brakke.[66] The potential energy for various displacements of the micropart relative to the binding site (see Fig. 3.40) and for different lubricant volumes was calculated. These displacements are schematically reproduced in Fig. 3.41.[68] They include a shift of the micropart against the binding site, a lift of the MEMS AND NEMS SIMULATION, KORVINK, RUDNYI, GREINER, LIU 169 CH03 9/9/05 8:50 AM Page 169 micropart in the direction perpendicular to the binding site, a twist rotation with the rotation axes taken as the z-axes, and a tilt motion, i.e., a rotation, with respect to an axis in the plane of the micropart. Different lubricant volumes were investigated. There is a clear dependence of the final alignment precision on the variations in the alignment forces and corresponding potential field"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002591_iccci.2013.6466157-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002591_iccci.2013.6466157-Figure1-1.png",
        "caption": "Figure 1. Free body diagram of the inverted pendulum system",
        "texts": [],
        "surrounding_texts": [
            "Keywords-inverted pendulum model; CLQR; DLQR\nI. NOMENCLATURE\n\u03b8 Pendulum deflection angle x Slider Position g Universal Gravitational Constant\npsM Mass of the Slider + Piston\nbM Mass of the Pendulum\n2L Length of the Stick\ncf Cylinder friction\nF Cylinder applied force \u03b6 Damping Ratio of the applied force model\nn\u03c9 Natural Frequency of the applied force model\nb Viscous Friction co-efficient\ndT Lead time costant of the applied force model\n0K Gain parameter oh the applied force model\npA Crossectional area of the piston\n0V Total air volume\n0P Atmospheric pressure\nsP Air supply pressure\n0\u03c1 Ambient air density\npK Proportional valve gain parameter\nP\u0394 Pressure difference between the chamber of the cylinder\ncF Coulomb friction\nsF Breakway friction\nsv Striebeck speed\n\u03b4 Striebeck co-efficient\nII. INTRODUCTION It is common practice to analyze the efficiency of the various control strategies in Inverted pendulum which is used as a non-linear system. Optimal control based on Linear Quadratic Regulator (LQR) method may be applied in a nonlinear system with the assumption of the initial states [1]. It is obvious to solve the Algebraic Riccati equation (ARE) that gives Riccati solution ( )P by choosing of appropriate weight matrix ( )Q and weight factor ( )R to design the linear quadratic regulator based on minimization of quadratic performance index ( )J as in [2]. This regulator may also be used as a state feedback gain ( )K that gives the control vector ( )u . For an Inverted pendulum system it is considered that initial angle very small ( )\u03b8 to design an optimal controller based on LQR technique in [3]. Chandan Kumar et al. [4] propose the Jacobian linearized model of an Inverted pendulum system to design the LQR based optimal controller. LQ and LQG based optimal controller design methodology to control a pneumatically actuated inverted pendulum system has been proposed in [5]. This paper deals with the linearized model of a non-linear pneumatically actuated inverted pendulum system [5]. The weighting matrices have been chosen to design CLQR and DLQR based optimal controller\n978-1-4673-2907-1/13/$31.00 \u00a92013 IEEE",
            "like as in [4] to minimize the performance index ( )J and compares both the results respectively.\nOn the other hand, K.J. Astrom and K. Furuta [6] have investigated some properties of the simple strategies energy control based swinging up the pendulum. Johnny Lam [7] has been implemented an energy controller and non-linear heuristic controller in order to swing the inverted pendulum to an upright position. An adaptive controller design technique based on Lyapunov stability theorem has been proposed for inverted pendulum system in [8]. Genetic algorithm (GA), Particle swarm optimization (PSO) and ant colony optimization (ACO) methods based Rotational Inverted Pendulum (RIP) controller has been designed to balance the inverted position of the pendulum in [9]. In [10], a variable-gain PID controllers and a sliding mode controller is deigned to regulate and control the pendulum arm respectively. Attitude control of a triple inverted pendulum has been proposed in [11].To gets the stability of the double-inverted pendulum at an upright position a fuzzy controller with the high-resolution and high-accuracy is designed in [12].\nOther significant work to design the optimal controller, an analytical approaches for guaranteed dominant pole placement, considering optimal sate feedback gains are the PID controller gains to design an inverse optimal controller in S. Das et al. [13]. Hongliang Wang et al. [14] has been analyzed the modeling and performance of a linear inverted pendulum. For the nonlinear inverted pendulum the Lyapunov optimal feedback control strategies have been discussed in W J Grantham et al.[15]. To solve the optimal control problem of nonlinear systems the least squares support vector machines (LS-SVM's) have been used by J.A.K. Suykens et al. [16]. A triple link inverted pendulum is controlled by a single-input feedback controller. The controller is designed by the use of non-linear optimization technique [17].\nRest of the paper is organized as follows: The model dynamics is described in section III. The optimal controller design reports in section IV. Simulation and studies are reported in section V. The paper ends with the conclusion in section VI, followed by the references.\nIII. MODEL DYNAMICS\nA. Dynamics of the Inverted pendulum The non-linear pneumatically actuated pendulum model [5] has been taken for the implementation of the proposed controller. The free body diagram of the system is shown in the figure (1). It has been assumed that the center of the gravity is located at the half length of the pendulum stick and the friction characteristic at the hinge point of the pivot is considered as negligible. Taking moment of inertia about the hinge point yields\n( )2 2 22 4\n12 3 b b b\nM L I M L M L= + = (1)\nThe pneumatic cylinder is acting as an actuator of the pendulum. The friction characteristics are characterized by pneumatic cylinder [5]. With the application of Lagrange's\nequation about the cylinder position ( x ) and the angular position (\u03b8 ) of the pendulum stick, gives the following nonlinear dynamical equations\n( ) ( )2cos sinb ps c psM M x f x M F\u03b8 \u03b8 \u03b8 \u03b8+ + + \u2212 = (2)\n24 sin cos 3 b ps psM L M Lg M Lx\u03b8 \u03b8 \u03b8= \u2212 (3)\nThe parameters are carrying their usual meanings as given in the nomenclature list.\nThe Dynamics of the rod less pneumatic cylinder which is driven by a proportion control valve has been taken from [5]. The air pressure dynamics are described as follows\n0 0 0 0/ 2 / 2 s p s p P A P PP x K u V V\u03c1 \u0394 = \u2212 + (4)\nThe dynamics of the proportional valve has been neglected and simply replaced by a gain factor, pK . This is because of the fact that the dynamics of the pressure development inside the cylinder is much slower than that of the dynamics of the proportional valve. The motion- position of the piston-slider in the cylinder has been realized by the following set of equations.\n( ) ( ) ( ) ( ) p\nps b ps b\nA P bx F bx x\nM M M M\n\u0394 \u2212 \u2212 = =\n+ + (5)\nNow equation (5) can be rewrite in terms of slider speed ( )x and substitute it into the equation (4) yields,\n( )2 2 02 n n n dF F F K T u u\u03b6\u03c9 \u03c9 \u03c9+ + = + (6)",
            "Where,\n( )\n( )\n( )\n2\n0\n0 2\n0 0\n0\n2\n2 2\ns p n\nb ps\ns p b ps\np p\nb ps d\nP A V M M\nVb P A M M\nPK K bA\nM M T\nb\n\u03c9\n\u03b6\n\u03c1\n\u23ab \u23aa=\n+ \u23aa \u23aa \u23aa\n= \u23aa\u23aa+ \u23ac \u23aa\n= \u23aa \u23aa \u23aa+ \u23aa= \u23aa\u23ad\n(7)\nThe characteristic of the friction is highly non-linear inside the cylinder. The behavior of the frictional force from the law of Striebeck static characteristic is given below as in [5],\n( ) ( ) ( )exp sgnc c s c s xf x F F F b x xv \u03b4\u23a1 \u23a4\u239b \u239e= + \u2212 \u2212 +\u239c \u239f\u23a2 \u23a5\u239d \u23a0\u23a3 \u23a6\n(8)\nB. State space equations of the system The non-linear state space representations [5] are considered as follows:\n[ ] [ ]1 2 3 4 5 6 TTx x x x x x x x x F F\u03b8 \u03b8\u23a1 \u23a4= =\u23a3 \u23a6 (9) Where x is the state vector, can be obtained from equation (1), (2) and (6) as follows:\n( )\n1 2\n2 3 3 4 3\n5 6 2 2\n2 2 3 3\n3 4\n23 3 5 4 3 2\n4 2 2 3 3\n5 6\n3 sin cos sin 4\n3 31 cos 1 cos 4 4\n3 sin 3cos sin4 cos cos3 341 1\n4 4\nb b d c\nps b ps b\nc\nb bb ps\nb ps b ps\nx x\nM g x x M Lx xx T x f xx M M x M M x\nx x\ng x x x x x f xLx M x M xL M M M M M M\nx x\nx\n=\n++ \u2212= \u2212 \u239b \u239e \u239b \u239e+ \u2212 + \u2212\u239c \u239f \u239c \u239f \u239d \u23a0 \u239d \u23a0\n=\n\u239b \u239e \u239c \u239f+ \u2212\u239c \u239f= \u2212 \u239c \u239f+\u2212 \u2212\u239c \u239f+ +\u239d \u23a0\n= 2 2\n6 5 6 02n n nx x K u\u03c9 \u03b6\u03c9 \u03c9 \u23ab \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23ac \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa= \u2212 \u2212 + \u23ad\n(10) So equation (10) can be represented as the form of ( , )x f x u= .\nThe equation (10) can be linearized by the assumption of the Inverted pendulum deflection angle \u03b8 is very small i.e.\n3 3 3 4sin ,cos 1, 0x x x x\u2248 \u2248 \u2248 and slider in the motion so that by the viscous friction model cylinder friction may be approximated in the form of 2cf bx bx= = . The linearized state space model can be articulated as [2],\n[ ]1 3 5 T\nx Gx Hu\ny Cx x x x\n= + \u23ab\u23aa \u23ac\n= = \u23aa\u23ad (11)\nWhere,\n0 1 0 0 0 0 3 44 40 0\n4 4 4 4\n0 0 0 1 0 0 3 ( )3 30 0 0\n4 4 4\n0 0 0 0 0 1 20 0 0 0 2\n20 0 0 0 0 0 1 0 0 0 0 0 0 0 1 0 0 0 0\nM g Tb b d M M M M M M M Mps b ps b ps b ps b\ng M MG b ps b\nL M M L M M L M Mps b ps b ps b\nn n T\nH K n\nC\n\u03c9 \u03b6\u03c9\n\u03c9\n\u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5\u2212 \u2212\u23a2 \u23a5+ + + + \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5+= \u23a2 \u23a5\u2212 \u23a2 \u23a5\u239b \u239e \u239b \u239e \u239b \u239e+ + +\u239c \u239f \u239c \u239f \u239c \u239f\u23a2 \u23a5\u239d \u23a0 \u239d \u23a0 \u239d \u23a0\u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\n\u2212 \u2212\u23a2 \u23a5\u23a3 \u23a6\n\u23a1 \u23a4= \u23a2 \u23a5\u23a3 \u23a6\n= 0 0 0 1 0 \u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5 \u23a2 \u23a5\u23a3 \u23a6\n(12)\nIV. OPTIMAL CONTROLLER DESIGN To design the LQR based optimal control, the system is controlled assuming to be a linear system and it must be controllable. This type of control approaches may be used for its good robustness properties.\nA. Continuous Linear Quadratic Regulator Yong Xin et al. [3] and Chandan Kumar et al. [4] have given an approach of design of optimal controller to control the angle of the inverted pendulum system on a cart. They have selected appropriate weighting matrices to get desired response from the system. As the LQR based state feedback controller is used for linear system, the system may be considered in terms of state space model as,\n( ) ( ) ( )x t Gx t Hu t= + (13)\nA Continuous Algebraic Riccati Equation (CARE) given in (14) can be solved with the help of weight matrix ( )Q and weight factor ( )R .\n1 0T TG P PG PHR H P Q\u2212+ \u2212 + = (14)\nWhere the weight matrix ( )Q is positive definite and weight factor ( )R is positive semi definite.\nThis Riccati solution ( )P which is obtained from the (14), is used to find out control vector ( )u that can be expressed as:\n1( ) ( ) ( )Tu t R H Px t Kx t\u2212= \u2212 = \u2212 (15)\nQuadratic performance index ( )J is minimized (16) which is expressed in the form of as in [2,18],\n0\n( ) ( ) ( ) ( )T TJ x t Qx t u t Ru t dt \u221e\n\u23a1 \u23a4= +\u23a3 \u23a6\u222b (16)\n( )x t is the state trajectory of the system."
        ]
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.45-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.45-1.png",
        "caption": "Fig. 3.45",
        "texts": [
            " The rotating body is then referred to as unbalanced. The moments perpendicular to the rotation axis are only zero if the products of inertia vanish, i.e. if the rotation axis is a principal axis. During dynamic balancing the products of inertia of the body are brought to zero (or nearly zero) by affixing additional masses to the rotating body. A body is called statically balanced if its center of mass is located on the rotation axis. E3.22Example 3.22 A thin homogeneous plate of mass m is free to rotate in supports A and B as shown in Fig. 3.45. Its rotation is driven by a constant moment M0. Formulate the equations of motion and determine the support reactions. Solution To describe the motion of the plate it would be sufficient to apply the principle of angular momentum about the fixed axis. However, to determine the support forces, the principles of angular momentum about the axes perpendicular to the rotation axis and the principle of linear momentum are needed. We free the plate from the supports and choose a body-fixed (rotating) coordinate system x, y, z whose origin A is at rest (fixed in space). The center of mass rotates on a circle (Fig. 3.45b). With centripetal acceleration acx = \u2212c \u03c92/3 and tangential acceleration acy = c \u03c9\u0307/3 the components of the principle of linear momentum in the x- and in y-direction read m acx = Ax + Bx \u2192 \u2212 mc \u03c92 3 = Ax + Bx , (a) m acy = Ay + By \u2192 mc \u03c9\u0307 3 = Ay + By . (b) To set up the components of the principle of angular momentum, the moments and products of inertia are needed. With dm = \u03c1 t dA and m = 1 2 \u03c1 t c b they are calculated as (cf. Fig. 3.45c) \u0398z = \u03c1 t \u222b x2 dA = \u03c1 t b\u222b 0 { cz/b\u222b 0 x2 dx } dz = mc2 6 , \u0398xz = \u2212 \u03c1 t \u222b xz dA = \u2212 \u03c1 t b\u222b 0 { cz/b\u222b 0 xdx } z dz = \u2212 mcb 4 , \u0398yz = 0 . Thus, from (3.61) we obtain Mx = \u0398xz \u03c9\u0307 \u2192 \u2212 b By = \u2212 \u03c9\u0307mcb 4 , (c) My = \u0398xz \u03c92 \u2192 b Bx = \u2212 \u03c92mcb 4 , (d) Mz = \u0398z \u03c9\u0307 \u2192 M0 = \u03c9\u0307mc2 6 . (e) The last equation represents the equation of motion. Assuming the initial condition \u03c9(0) = 0 leads to 3.4 Kinetics of a Rigid Body in Three Dimensional Motion 197 \u03c9\u0307 = 6 M0 mc2 \u2192 \u03c9 = 6 M0 mc2 t . Insertion into (a) - (d) finally yields Ax = \u2212 \u03c92mc 12 , Ay = M0 2 c , Bx = \u2212 \u03c92mc 4 , By = 3 M0 2 c "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003801_j.ymssp.2013.03.014-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003801_j.ymssp.2013.03.014-Figure2-1.png",
        "caption": "Fig. 2. The magnetic system.",
        "texts": [
            " However, if there is a nonlinear component inside the system and it is not feasible to linearize the nonlinearity, then the whole system has to be considered as a nonlinear one and the nonlinear control design needs to be employed. In the previous section, we have introduced a method to compensate an uncertain nonlinear block into a known linear mapping. Hence, if both the input and output of the nonlinearity are available then a feedback path can be established to cover its nonlinearity and uncertainty so that a linear mapping is obtained. Therefore, the linear control theory can still be feasible to stabilize the whole system. To verify the proposed method, a set of experiments are conducted on a magnetic system (Fig. 2) [18]. The input to the system is the voltage v across the coil which generates the magnetic force F via the current i in the coil. The mapping from v to i is well-known to contain a hysteresis [5,8] and can be described by the relationship di dt \u00bc v\u2212iR\u00fe \u2202\u03bb\u00f0x,i\u00de \u2202x dx dt \u2202\u03bb\u00f0x,i\u00de \u2202i \u22121 \u00f018\u00de where R is the reluctance of the coil, x is the displacement of the moving mass, and \u03bb is the flux linkage. The dynamics in (18) can be regarded as the relationship i\u00bc h\u00f0v\u00de which is exactly the same as the one in Fig",
            " Here, we would like to present the rate-dependent hysteretic behavior in this mapping experimentally by observing the output current when the input voltage is varying with different frequencies. The proposed compensator is then applied to modify the dynamics into a linear mapping. In addition, we are going to integrate this design into a control system so that a linear controller is sufficient to give good performance. The magnetic force F generated by the magnet is applied to the mover represented as the mass m in Fig. 2. The mover is able to move on a linear guideway MGN-9H by Hiwin Technology. Its displacement is measured by an encoder from Micro E System with the type Mercury 35100 SS-350c. The encoder pulses are fetched by the HCTL-2032 chip which gives 1 \u03bcm accuracy. The current in the coil is measured via a current sensor by LEM. It is connected to the A/D channel of the AX5412-H card by Axiomtek. The control strategy is implemented in the timer interrupt service routine of a PC-compatible to have a 0.5 ms sampling period",
            " The current error converges very fast which can be seen in Fig. 6(b) where only the first 0.2 s are shown. The steady state current error is within 0.02 A. Fig. 7(a) gives the input voltage u and 7(b) is the compensation effort \u03b7. Fig. 8 presents the input/output relationship after compensation under various input frequencies. It can be seen that the relationship is almost linear regardless of the uncertainties and rate-dependent nonlinearities. Compared with Fig. 5, this is a significant improvement. The magnetic system in Fig. 2 can be used as a position servo system. The equation of motion of the mechanical part is expressed as d2x dt2 \u00bc 1 m \u00f0\u2212F\u00femg\u00de \u00f019\u00de i.e., the mass m is driven by the magnetic force F under the effect of the gravity. Therefore, the system is open-loop unstable, and a stabilization controller is needed for proper manipulation of the mass. The magnetic force F is generated by the current i via a known static mapping \u03c6\u00f0i\u00de [19] which can be easily obtained analytically or experimentally. This way we may represent the system in a block diagram as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003321_s12206-011-0630-6-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003321_s12206-011-0630-6-Figure4-1.png",
        "caption": "Fig. 4. String bed model for a parametric study.",
        "texts": [
            " The displacement vector, [ ,1]IJ k y , is computed by multiplying the inverse of the linearized structural stiffness matrix with the external transverse force vector [ ,1]IJ k F . The internal transverse force vector, [ ,1]( ) IJ k pF is determined by the multiplication of the structural stiffness matrix K with the displacement vector, [ ,1]IJ k y . The norm of force difference between the external transverse force vector [ ,1]IJ k F and the internal transverse force vector [ ,1]( ) IJ k pF is compared with a small number \u03b5 to see if the external transverse force vector [ ,1]IJ k F and the internal transverse force vector, [ ,1]( ) IJ k pF , are in equilibrium. Fig. 4 shows a schematic of the head in a tennis racket with strings strung. The strings strung vertically and the strings strung horizontally are called main strings and cross strings, respectively. The head can be approximately represented by an elliptic shape with horizontal radius, ,a and vertical radius, .b Main strings and cross strings are considered to fill up the entire head as much as they can. While the distance along the horizontal radius between the outermost main string and the edge of the head is defined as 0 ,a the distance along the vertical radius between the outermost cross string and the edge of the head is defined as 0 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001230_0470036427.ch3-Figure3.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001230_0470036427.ch3-Figure3.11-1.png",
        "caption": "Figure 3.11 Power as the product of voltage and current, with current lagging behind voltage by a phase angle f.",
        "texts": [
            " The average power (the average product of voltage and current) can be obtained by taking the averages (rms values) of each and then multiplying them together. Thus, Pave \u00bc IrmsVrms (resistive case) Power for the resistive case is illustrated in Figure 3.10. But now consider a load with reactance. The relative timing of voltage and current has been shifted; their maxima no longer coincide. In fact, one quantity is sometimes negative when the other is positive. As a result, the instantaneous power transmitted or consumed (the product of voltage and current) is sometimes negative. This is shown in Figure 3.11. We can interpret the negative instantaneous power as saying that power flows \u201cbackwards\u201d along the transmission line, or out of the load and back into the generator. The energy that is being transferred back and forth belongs to the electric or magnetic fields within these loads and generators. Since instantaneous power is sometimes negative, the average power is clearly less than it was in the resistive case. But just how much less? Fortunately, this is very easy to determine: the average power is directly related to the amount of phase shift between voltage and current",
            " By convention, a lagging current has a negative phase angle (as compared to the voltage, which is taken to have a zero phase angle), and the resulting power has a positive phase angle: S \u00bc Irms/ u Vrms/08 \u00bc Irms/u Vrms/08 \u00bc IrmsVrms/u Thus, the phase angle u of complex power S is the same as the angle by which current is lagging voltage. This is the same result we claimed without further justification in Section 3.3.2, where the angle f in the triangle in Figure 3.12 is the same as the phase angle f between the sine waves in Figure 3.11. Because complex, real, and reactive power all have the same physical dimension (energy per time), despite their different labels (VA, watts, and VARs) for identification purposes, they can usefully be combined in a phasor diagram. Indeed, Figure 3.12 is precisely such a diagram, except that, when introducing it, we did not mention any implicit rotation of the arrows, nor is it necessary to retain the visual image of rotation in order to apply the diagram for calculation purposes in practice. Considering the importance, simplicity, and explanatory richness of that very basic drawing, we can appreciate why phasors are an electrical engineer\u2019s most essential tool for visualizing a"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure2-1.png",
        "caption": "Fig. 2. PAMD Ideal Portable View",
        "texts": [],
        "surrounding_texts": [
            "The detailed CAD designs of the PAMD using the NX 7.5 software are shown in Figures 1 and 2 with dimensions."
        ]
    },
    {
        "image_filename": "designv11_100_0002886_anthology.2013.6784715-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002886_anthology.2013.6784715-Figure1-1.png",
        "caption": "Figure 1. Bicycle model of a vehicle.",
        "texts": [
            " In this new RRTs planner we make use of user\u2019s capacity of perception and computer\u2019s capacity of autonomous compute, combine human intelligent and computer intelligent, giving attention to both humancomputer interactions and dynamic constrains of robots to yield much better performance in complex environments than autonomous RRTs planners. We validate the results in lots of experiments. II. ROBOT DYNAMIC MODEL Considering dynamic constrains of the mobile robot in path planning, we must have a dynamic model. The normal simplified robot model which is called \u201cbicycle model\u201d and presented here is based on [15]. We use a linear tire model described in [16]. Fig. 1 shows the top view of the vehicle using this bicycle model. The non-linear bicycle model considers longitudinal ( xv ), lateral ( ), and yaw (r) motion under the assumption of negligible lateral weight shift, roll and compliance steer while traveling on a smooth road. In order to simplify the model and make it convenient in planning, we set a constant longitudinal speed yv xv for this model. In addition, longitudinal and lateral positions and yaw angle ( , , )X Y with respect to the fixed inertial coordinates are added to the dynamic equation in order to refresh the vehicle position and orientation in the scene of RRT growing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.12-1.png",
        "caption": "Figure 11.12: Instrument panel (revised design) CAE model.",
        "texts": [
            " In the revised IP, the inner wall of the glove box door is reinforced with ribs with the shape of an \u201cegg crate\u201d, which is designed to absorb fore-aft vibrations critical to squeak and rattle performance. The design concept aims at not only minimizing squeak and rattle concerns but also enabling designers to reduce glove box door fasteners from six to merely two. The revised steering column has a positive clamp tilt lever mechanism, which significantly increases the lateral stiffness of the column system. The CAE model of this IP is shown in Figure 11.12. Accelerations of 72 fasteners in this IP were computed using a vehicle CAE model and the peak response of each fastener acceleration was identified. An examination of all response peaks indicated that none of the fasteners has an acceleration higher than 10,000 mm/sec2. The fastener acceleration at the glove box door was reduced significantly as shown in Figure 11.13. The single fastener with the highest acceleration occurs at the central cowl top as shown in Figure 11.14. The acceleration response of this fastener is given in Figure 11"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003439_1.4771592-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003439_1.4771592-Figure1-1.png",
        "caption": "FIG. 1. Three possible uniaxial nematic phases (N1, N2, N3) and two biaxial nematic phases (N1b, N2b) of side-chain liquid crystalline polymers. Depending on two orientational order parameters: one is that Sm of a sidechain (mesogen) and the other is Sb of a rigid-backbone chain, we can define a nematic N1 phase with Sm > 0 and Sb < 0, a N2 phase with Sm < 0 and Sb > 0, and a N3 phase with Sm > 0 and Sb > 0. Biaxial nematic N1b phase is defined as a minor director db appears perpendicular to the major director Dm. Biaxial nematic N2b phase is defined as a minor director dm appears perpendicular to a major director Db.",
        "texts": [
            "25 Biaxial phases of side-chain liquid crystalline polymers (LCPs) have been studied in nematic and smectic A phases with sideon or end-on attachment of mesogenic group to the polymer backbone.26\u201330 When we consider a nematic phase of side-chain LCPs, there are three uniaxial nematic phases, which were first considered by Wang and Warner.31 The three nematic phases can be defined by two orientational order parameters: one is Sm of nematogenic side-chains (mesogens) and the other is Sb of a rigid-backbone chain. When one order parameter is positive, the other can be positive or negative.32, 33 Figure 1 shows three principal uniaxial nematic phases for a side-chain LCP. The N1 phase (Sm > 0 and Sb < 0) is defined as that the mesogens are aligned along to the ordering direction (z) and the backbone chains are randomly distributed on the plane perpendicular to the director Dm. The N2 phase (Sm < 0 and Sb > 0) is defined as the backbone chain is aligned along to the ordering direction Db and the mesogens are randomly distributed on the plane perpendicular to the backbone chain. The third N3 phase is defined by Sm > 0 and Sb > 0, where the backbone and mesogens are oriented along to the ordering direction Dmb. In the N1 phase, the backbone chain adopts an oblate shape. In the N2 and N3 phases, a prolate shape of the backbone is obtained.34\u201340 In these uniaxial nematic phases (N1, N2), we can expect biaxial nematic phases. Figure 1 schematically shows novel biaxial nematic phases (N1b, N2b) of side-chain LCPs, where the mesogens and backbone chains favor mutually perpendic- a)Electronic mail: matuyama@bio.kyutech.ac.jp. URL: http://iona.bio. kyutech.ac.jp/~aki/. ular orientations.41 In the uniaxial nematic N1 phase, we can expect additional ordering of backbone chains in the direction db (minor director) perpendicular to the director Dm (major director). This corresponds to biaxial ordering (N1b). In the N2 phase, we may have a biaxial nematic (N2b) phase, where the additional ordering of mesogens appears in the direction dm (minor director) perpendicular to the major director Db"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003717_amr.383-390.2622-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003717_amr.383-390.2622-Figure1-1.png",
        "caption": "Fig. 1.Bearing structure",
        "texts": [
            " Bearing failure was the second only to gear in all faults in the gearbox. Bearing condition monitoring and fault diagnosis not only play a vital role in understanding the performance state and early detection of potential failure, but also can improve the operation and management level of machinery and equipment and maintenance performance, which has significant economic benefits. Structure of Rolling Bearing. The basic structure of rolling bearing consists of the four parts, for example inner ring, outer ring, ball and cage, as shown in Fig. 1. The inner ring is installed in the journal, outer ring is installed in the bearing hole, the inner usually rotates with axis and the outer ring fixes. There is also the outer ring of bearing rotated, inner ring fixed, such as the bearing on the front wheel of automotive. And there is inner and outer rings rotated simultaneously at different speeds, such as the bearing on planetary axle. Rolling body is the core component of rolling bearings. When the inner and outer rings relative rotates, the balls run in the raceway between the inner and outer rings",
            " The shapes of the rolling body usually include ball, cylindrical roller, needle 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 TTP, www.ttp.net. (ID: 130.113.86.233, McMaster University, Hamilton, Canada-24/03/15,18:39:31) the mostly-shaped groove, which can reduce the contact stress and limit the axial movement. Roll cage separate rolling body evenly in order to avoid direct contact between the rolling bodies and reduce heat and wear. Cage includes pressed cage (Fig.1.a) and solid cage (Fig.1.b). The Common Fault of Bearing. Bearing is the vulnerable machine part and a common mechanical component which is widely used in variety of rotating machinery. Serious bearing failure will lead to severe vibration and noise in the machine and reduce equipment efficiency, and even cause equipment damage. Every point defect on bearing, such as rolling body fatigue erosion, raceway wear, the cage deformation or breaking, the friction which caused from inside and outside the ring loose with the hole or axis, and lack of oil or oil mixed with impurities, will be reflected in the bearing vibration signals"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002268_kem.464.358-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002268_kem.464.358-Figure4-1.png",
        "caption": "Fig. 4 Third order of vibrated mode Fig. 5 Forth order of vibrated mode",
        "texts": [],
        "surrounding_texts": [
            "Model Analysis. Some parameters are set up before simulation. The unit type is Solid92. Elastic modulus is 2.06E11 2kg/(m s )\u22c5 . Material Poisson's ratio is 0.3. Material density is 7800 3kg/m . Radial stiffness is set up 2.2E7N/m. Unit size is 0.005m. The simulation can obtain front 4 steps natural frequency and mode shapes as follows: From the above figure2-5, we can get nature frequency and vibration mode of LRG. Nature frequency and vibration direction are details in table.1. During the analysis of mode, in order to get the regular pattern between spring rate and natural frequency, this paper takes second-order vibration mode as example. Spring rate is variable, and the multi-groups datum are gathered to carry on the comparison, and then matched curve by method of non-linearity least squares in MATLAB software simulation. Figure.6 shows that stiffness value increases, natural frequency correspondingly increases. With the stiffness continuing increase, the natural frequency gradually tends to stability. Fig.6 Relationship between natural frequency and joint stiffness"
        ]
    },
    {
        "image_filename": "designv11_100_0003445_ht2012-58139-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003445_ht2012-58139-Figure9-1.png",
        "caption": "Fig. 9: Deformation along weld line and comparison with undeformed workpiece. Distortions emphasized by factor of 100.",
        "texts": [
            " The results highlighted 7 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/01/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use here are for the parameters: welding current 150A, welding speed 2mm/s and arc length 3mm. After the butt joint is welded together, the arc is extinguished, and the workpiece cools down to ambient levels by dissipating heat into the surroundings. All deformation/stress results discussed here are after the temperature of the workpiece reaches ambient conditions. Figure 9 shows the contours of the total deformation along the weld line and the total workpiece distortion. The un-deformed workpiece is also superimposed in the figure. It is to be noted that to emphasize the modeled deformation of the structure after welding, the distortions are multiplied by a factor of 100. Figure 10 shows the total deformation (on a true scale) of the thin plate after welding. The welded thin plate undergoes distortion due to thermal gradients. With reference to Fig. 5, we find that surface A\u2019B\u2019C\u2019D\u2019 was clamped (fixed) and hence shows no deformation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.36-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.36-1.png",
        "caption": "Fig. 2.36 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PaPRP (a) and 4PaRPP (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology Pa\\P||R\\P (a) and Pa\\R\\P\\kP (b)",
        "texts": [
            "32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.1-1.png",
        "caption": "Fig. 3.1 A clock pendulum as an idealised compound pendulum. (a) Rest position. (b) Deflected",
        "texts": [
            " Except for idealisations of compound pendulums, these idealisations are all assembled from components of simple rod pendulums or simple string pendulums (Sect. 2.4). Most can display chaotic behaviour. An idealised compound pendulum is a rigid body suspended from a frictionless horizontal pivot. Rigid means that the body does not deform under an applied load. It is assumed that the only forces acting are those due to gravity, acting vertically downwards, on the component parts of the body. A clock pendulum as an idealised compound pendulum is shown in Fig. 3.1. L.P. Pook, Understanding Pendulums: A Brief Introduction, History of Mechanism and Machine Science 12, DOI 10.1007/978-94-007-1415-1 3, \u00a9 Springer Science+Business Media B.V. 2011 27 For a body that can rotate about an axis, if r is the perpendicular distance from the axis of an element mass m, then the summation P mr2 is the moment of inertia, I , of the body about the axis. For simple shapes values of I can be obtained by integration and are listed in textbooks, for example Loney (1913). If M is the mass of the body and taking M 2 D X mr2 (3.1) then is the radius of gyration of the body about the axis. For a simple rod pendulum (Fig. 2.1a) the moment of inertia of a is ml2 where m is the point mass and l is the length of the rod, and the radius of gyration about the axis (frictionless pivot) is l . For small amplitudes (value of in Fig. 3.1b) the derivation of the pendulum period, P , follows that for the simple rod pendulum given in Sect. 2.3 (Loney 1913; Lamb 1923). The moment about the frictionless pivot is Mghsin where g is the acceleration due to gravity, h is the distance of the centre of mass, often called the centre of gravity, from the frictionless pivot, and is the pendulum angle. Hence I d2 dt2 D Mghsin (3.2) where I is the moment of inertia, and t is time. This is the same as for a simple rod pendulum if l D I=Mh D 2=h (3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003323_ijtc2011-61263-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003323_ijtc2011-61263-Figure3-1.png",
        "caption": "Fig. 3 A cracked tooth alters gear-transmitted torque, generating noise, difference between actual and ideal signals.",
        "texts": [
            "org/about-asme/terms-of-use 2 Copyright \u00a9 2011 by ASME angle and movement of the contact over the face of the tooth; tooth contact mechanics, including involute profile, normal and frictional forces, Hertzian contact stiffness between contacting teeth, and slip and roll kinematics of spur gear contacts; and finally, multiple tooth contact, including contact ratio. By adjusting certain parameters, simulations can overlay the complex signals measured by sensors. Tuning parameters and following changes can detect and track faults, since the model was built with direct physical correspondence between parameters and specific components (and faults). For example, a cracked gear tooth with larger bending compliance perturbs the torque transmitted across gears, as shown in the Fig. 3. The uncracked tooth compliance can be estimated by beam theory applied to a tapered tooth, or by tuning to data for healthy gears. Parameter tuning can be automated, or manual. To quantify the degraded functional capability of the gearbox due to faults, treat the gearbox in Fig. 1 as a communications channel, where input u(t) \u201ctransmitted\u201d over the \u201cmachine channel\u201d is \u201creceived\u201d as output y(t). Faults along the gearbox that disrupt usual flow of signal add \u201cnoise\u201d n(t) = y(t) - yi(t), \u201c\u2026any unwanted component in a received signal\u201d [5], tantamount to the difference between the signal yi(t) that would be received with no faults, and actual output y(t)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002666_cdc.2012.6426701-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002666_cdc.2012.6426701-Figure1-1.png",
        "caption": "Fig. 1. The satellite-rotor system",
        "texts": [
            " In this section we overview the rotational dynamics of a spacecraft with a rotor. In particular, we use the matrix representation of angular velocity and momentum, which is 978-1-4673-2064-1/12/$31.00 \u00a92012 IEEE 1285978-1-4673-2066-5/12/$31.00 2012 I a step towards the transition from the continuous-time to discrete model. Following Krishnaprasad [24] and Bloch, Krishnaprasad, Marsden, and Sa\u0301nchez de Alvarez [4], we consider a rigid body with a rotor aligned along the third axis of the body as shown in figure 1. The rotor spins under the influence of a control torque u. The configuration space is Q = SO(3)\u00d7SO(2), with the first and second factors being the configuration spaces of the spacecraft and rotor, respectively. Let g \u2208 SO(3) be the attitude of the system, \u2126 = g\u22121g\u0307 \u2208 so(3) be the angular velocity of the system measured against a body frame, \u03b7 \u2208 R be the angular velocity of the rotor relative to the spacecraft, and I, J : so(3) \u2192 so\u2217(3) be the inertia operators of the spacecraft and the rotor, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002009_978-3-319-01228-5_4-Figure4.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002009_978-3-319-01228-5_4-Figure4.7-1.png",
        "caption": "Fig. 4.7 Schematic representation of a manipulator with joint elasticity",
        "texts": [
            " Joint elasticity is a common problem in manipulators and represents the most important type of underactuated multibody systems with non-collocated output. Elasticities can occur in the transmission and reduction elements between the motors and the links of the manipulator. Those elements are for example harmonic drives, gears, long shafts or belts. These elements introduce a significant amount of elasticity, which must be considered in the model and control design. Thereby, the elasticity is lumped and modeled as a torsional spring with stiffness c between the motor and link. A schematic representation is given in Fig. 4.7 for a one link manipulator, where the generalized coordinates qm, ql describe the motion of the motor and link, respectively. The motor torque u acts directly on the motor variable qm . In addition to the elasticity also damping can by introduced by a parallel torsional damper. This model of a manipulator with joint elasticity has twice as many degrees of freedom than a traditional rigid manipulator. Since the end-effector position is determined by the generalized coordinates ql of the links, these unactuated coordinates must be tracked in the control problem"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002970_amm.86.619-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002970_amm.86.619-Figure2-1.png",
        "caption": "Fig. 2 Force figure of housing",
        "texts": [
            " In static force analysis inside housing, how to select loading and regulation methods is of vital importance. Only a right selection can ensure a most proximity to simulated actual working conditions. A three-dimensional model is built (as shown in Fig. 1). Force analysis has two steps. Step 1: Cohesive joints are selected and their rigidity matrix is worked with ANSYS. Step 2: The matrix is introduced into MASTA to work out the results of forces on each cohesive joint from loads, as shown in Fig. 2 below. Housing is gridded into 8-joint-and-6- surface units with FEA. It follows 2 steps. Step 1: general gridding and calculation. Step 2: detailed gridding (as shown in Fig. 3) based on force situation. The stronger forces are, the closer grids are. The final results (as shown in Fig. 4-a & Fig. 4-b) have been calculated: unit total \u2013 206,149; joint total \u2013 277,989. As calculated, the maximum equivalent stress on housing (as shown in Fig. 4-a) is 104MPa and the maxim distortion by comprehensive displacement (as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002378_s11232-013-0053-x-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002378_s11232-013-0053-x-Figure4-1.png",
        "caption": "Fig. 4. The Schwarz trick for straightening trajectories for standard billiards: the bold line indicates the billiard boundary to which a ray is incident.",
        "texts": [
            " The particular case V\u0303\u2016n = 0 of this regime corresponds to a stable composite particle (whose center of mass is at rest) with periodically changing velocities of the components. The case where the envelope charge is concentrated on one of its faces is the most difficult, and we consider it below after discussing the generalization of the Schwarz principle. One of most effective and widely used methods for studying the motion of rays in billiards was proposed by Schwarz. It is based on the idea of straightening a billiard trajectory [10] or on passing from a motion in a billiard to that of a ray in an infinite space without boundaries (see Fig. 4). This ideology is convenient when considering the motion of a composite particle in a constant electric field, but merely reflecting the billiard with respect to the boundary to which the trajectory is incident does not suffice in this case: we must generalize the Schwarz trick by appropriately changing the electric field in the reflected billiard image. Correctly continuing the trajectory in the reflected billiard therefore implies changing the electric field potential appropriately. It is easy to understand that together with reflecting the billiard with respect to the boundary on which the charged particle is incident, we must also reflect the electric field (see Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002982_amr.317-319.281-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002982_amr.317-319.281-Figure4-1.png",
        "caption": "Fig. 4. Finite element model of harmonic drive system",
        "texts": [
            " Therefore, it is not suitable to choose 1/4 or 1/2 of the flexspline model to investigate the transient dynamic problem. Instead, the whole model of the flexspline is to be applied in this work. Modeling method To obtain more accurate results, we use 8-node hexahedral element to model flexspline, circular spline and wave generator. Different parts adopt different mesh density. The areas of fierce stress or the regions of large stress grads are divided denser than those with relatively gentle stress. The assembled finite element model of harmonic gears drive is shown in Fig.4. Material parameters of harmonic gear drive are listed in Tab.1. To effectively solve the problem when the relatively large contact object surface penetrates the other object surface, we use automatic face-to-face contact mode to simulate mesh states between wave generator and flexspline, and between circular spline and flexspline. In the LS-DYNA, the element material properties defined by*MAT_RIGID keywords are used in some elements to accurately simulate the function of bearing to transmit load",
            " The solid elements of wave generator are added respectively with two layers of shell elements to simulate the rotational speed and limit the displacement boundary of the wave generator. The solid elements of output shaft are also added with two layers of rigid shell elements to simulate the moment on the output shaft. In the harmonic gear drive, there are some contact surfaces between wave generator and flexspline, and between circular spline and flexspline. Since solving the problem is computationally intensive, we define two contact regions in Fig.4 to reduce the contact search time and neglect the effects of wave generator bearing balls in the model to increase the compute speed. The transient dynamic analysis is carried out using the model shown in Fig. 4 under the 1000rpm rotational speed and 25N\u00b7m load, on the basis of large displacement and large strain theory. In the Fig. 3(b), sections of flexspline start from flexspline cup rim along the axial direction, having the lengths of 0.04, 0.16, 0.24, 0.27, 0.28 and 0.3 times of flexspline diameter, respectively. The stress variation curves on these sections are shown in Fig.5. Tab. 1 Material parameters of harmonic gear drive Components of harmonic gear drive Flexspline Circular spline, wave generator and output shaft Material type 30CrMnSi 45# Steel Tensile modulus [GPa] 204 210 Poisson's ratio \u00b5 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.8-1.png",
        "caption": "Fig. 2.8 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PRRRR (a) and 4RPRRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology P||R||R||R||R\\R (a) and R||P||R||R\\R (b)",
        "texts": [
            " 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7. 4RRRPR (Fig. 2.5a) R||R\\R\\P\\kR (Fig. 2.1f) Idem No. 5 8. 4RRPRR (Fig. 2.5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10. 4PRRRR (Fig. 2.6b) P||R\\R||R\\R (Fig. 2.1i) Idem No. 9 11. 4RRPRR (Fig. 2.7a) R\\R\\P\\kR\\R (Fig. 2.1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2.8a) P||R||R||R||R\\R (Fig. 2.1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig. 2.9a) R||R||P||R\\R (Fig. 2.1n) Idem No. 13 16. 4RRRPR (Fig. 2.9b) R||R||R||P\\R (Fig. 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No. 13 18. 4RRPRR (Fig. 2.10b) R||R||P||R\\R (Fig. 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.18-1.png",
        "caption": "FIGURE 1.18. Two samples of wire spoke wheel.",
        "texts": [
            "3 The ange shape code signi es the tire-side pro le of the rim and can be , , , , , , , , , and . Usually the pro le code follows the nominal rim width but di erent arrangements are also used. Figure 1.17 illustrates how a wheel is attached to the spindle axle. Example 37 Wire spoke wheel. A rim that uses wires to connect the center part to the exterior ange is called a wire spoke wheel, or simply a wire wheel. The wires are called spokes. This type of wheel is usually used on classic vehicles. The highpower cars do not use wire wheels because of safety. Figure 1.18 depicts two examples of wire spoke wheels. Example 38 Light alloy rim material. Metal is the main material for manufacturing rims, however, new composite materials are also used for rims. Composite material rims are usually thermoplastic resin with glass ber reinforcement, developed mainly for low weight. Their strength and heat resistance still need improvement before being a proper substitute for metallic rims. Other than steel and composite materials, light alloys such as aluminum, magnesium, and titanium are used for manufacturing rims"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.20-1.png",
        "caption": "Fig. 2.20 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRCR (a) and 4RCRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R||C||R (a) and R||C\\R||R (b)",
        "texts": [
            " 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig. 2.1g0) Idem No. 13 34. 4RRRC (Fig. 2.19a) R\\R||R||C (Fig. 2.1h0) Idem No. 21 35. 4RCRR (Fig. 2.19b) R\\C||R||R (Fig. 2.1l0) Idem No. 21 36. 4RRCR (Fig. 2.20a) R\\R||C||R (Fig. 2.1b0) Idem No. 21 37. 4RCRR (Fig. 2.20b) R||C\\R||R (Fig. 2.1c0) Idem No. 21 j\u00bc1 fj 5 5 23. Pp2 j\u00bc1 fj 5 5 (continued) In the fully-parallel topologies of PMs with coupled Sch\u00f6nflies motions F / G1G2-G3-G4 presented in this section, the moving platform n : nGi (i = 1, 2, 3, 4) is connected to the reference platform 1 : 1Gi : 0 by four spatial complex limbs with four or five degrees of connectivity. The complex limbs combine only revolute, prismatic and cylindrical joints. One actuator is combined in each limb. The actuated joint is underlined in the structural graph"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003226_icuas.2013.6564716-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003226_icuas.2013.6564716-Figure2-1.png",
        "caption": "Fig. 2. A coordinate system of the CNU-ducted fan UAV",
        "texts": [
            " A configuration of the CNU-ducted fan UAV is conventional ring-wing type as shown in Fig. 1. It has four control surfaces that are located at the end of the duct. Also, it contains fixed stators for anti-torque and additional lift. A fuselage is in center of the vehicle, and avionics is mounted in the duct or the fuselage. In addition, payload bay is placed on top of the fuselage. For various missions, operating equipments such as camera, spot light and communication relay can be located at this bay. A coordinate system shown in Fig. 2 has dynamic features similar to a helicopter: thrust vector, anti-torque, gyroscopic coupling, and velocity induced by a main rotor. Pitch angle and angle of attack are zero at hover flight: as the vehicle goes forward, it becomes negative[2]. In addition, moment of inertia is completely same about x- and y- axes because of symmetricity. The control surfaces are defined in Fig. 3: 1 and 3 are ailerons, 2 and 4 are elevators, and deflecting all of control surfaces are rudders. These deflect from -30 to +30 degree"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003543_20120215-3-at-3016.00205-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003543_20120215-3-at-3016.00205-Figure1-1.png",
        "caption": "Fig. 1. Test bench",
        "texts": [
            " The TMM is an analytical technique to calculate the eigenfrequencies and eigenmodes for arbitrary beam systems with diverse boundary conditions, see Pestel, C., and Leckie, F.A. (1963). In this paper the eigenfrequencies and eigenmodes are calculated for free-free beams, implying the shear force and bending moment to be zero at the boundaries of the beams. The benefit using the free-free modes as shape functions is that no moving direction is restricted and the degree of freedom is maximized. The test bench is shown in Fig. 1. It consists of a circular shaft, two discs including asymmetric masses and a motor for the acceleration of the shaft. The mounting is realized by two self-aligning ball bearings. This type of bearing has the advantage to be insensitive to angular misalignments of the shaft relative to the housing. A simple tooth belt realizes the connection between the motor and the rotor shaft. This type of mounting was chosen to minimize the influence of parameter excited vibrations which could occure when using a gear system for the coupling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.112-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.112-1.png",
        "caption": "Fig. 2.112 4CPaRR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology C\\Pa\\kR\\R",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure4-1.png",
        "caption": "Figure 4. Structure of ball Joint.",
        "texts": [
            " Steel press center structure is a strong combination because flange is fixed horizontal direction as shown in Figure 3. So the structure of steel and plastic were fixed in any direction. In general, ball joints that deliver the force from the tires are important parts with higher strength and durability. Also because ball stud often causes oscillation and rotation, ball joints bearings must be strong structure to withstand the steady force. Ball stud and steel pipes are fixed to the mold and plastic is injected into the mold in order to increase the strength and stiffness for ball joint part such as a Figure 4. For ball stud in order to smooth rotate and oscillate, plastic and bearings should be well fixed. However, because using the same materials of plastic and bearing, do not be combined with each other structure. So plastic and bearing slip will occur. To prevent slip for vertical and horizontal rotation, upper and lower bearings surface were applied to the \u2018 \u2019 shape. And to combine ball stud, upper and lower bearings were designed as a locking structure as shown in Figure 5. Suspension bush is like the joints for human"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003445_ht2012-58139-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003445_ht2012-58139-Figure10-1.png",
        "caption": "Fig. 10: Total deformation of the welded butt joint after completion of welding.",
        "texts": [
            " After the butt joint is welded together, the arc is extinguished, and the workpiece cools down to ambient levels by dissipating heat into the surroundings. All deformation/stress results discussed here are after the temperature of the workpiece reaches ambient conditions. Figure 9 shows the contours of the total deformation along the weld line and the total workpiece distortion. The un-deformed workpiece is also superimposed in the figure. It is to be noted that to emphasize the modeled deformation of the structure after welding, the distortions are multiplied by a factor of 100. Figure 10 shows the total deformation (on a true scale) of the thin plate after welding. The welded thin plate undergoes distortion due to thermal gradients. With reference to Fig. 5, we find that surface A\u2019B\u2019C\u2019D\u2019 was clamped (fixed) and hence shows no deformation. Greater deformation is observed on moving closer to the welding line, with maximum deformation at the beginning and end of the welding path. This can be explained on the basis of end-effect. The plots for directional deformation are not shown here, but it is seen that the symmetry face ABCD shows no deformation in the x-direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003653_iceee.2012.6421204-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003653_iceee.2012.6421204-Figure1-1.png",
        "caption": "Figure 1. Prototype \"La Raie Manta\"",
        "texts": [],
        "surrounding_texts": [
            "1) Frames: We assume the Earth is regarded as at and stationary in inertial space and de ne the inertial frame I = xI ; yI ; zI is a frame attached to the Earth. The body frame B = xB ; yB ; zB is tied to the center of gravity (CG) of the UAV and taking the xB axis xed to a longitudinal reference line in the simetry plane of th airplane. We de ne the stability frame A = xA; yA; zA as a frame xed to the aerodynamic center of the wing. The angular position of this frame depends on the angle of attack B of the UAV with respect to the relative direction of the wind. There is also a frame called wind frame S = xS ; yS ; zS whose origin is also at the aerodynamical center of the wing, but the angular position of this frame depends on both the angle of attack and sideslip angle of the UAV. These two last frames refer exclusively to the aerodynamic effects of the vehicle [2], [3]. Finally we have the mobile frame M = xM ; yM ; zM which is placed on an axis parallel to the yB axis and it passes through the center of mass of the frontal motors. 2) General vectors and rotational matrix: We de ned the general vector Q = [q1; q2] 2 R 6 where q1 = [x; y; z] T 2 R3 represents the coordinates of spatial position relative to the inertial frame I . The angular coordinates of the vehicle are represented by the vector q2 = [ ; ; ] T 2 R3 which are the Euler's angles [1],[2] . For the translational dynamics the general vector of forces is de ned in the body frame: FB = FBA + F B M T where FBA is the vector of the aerodynamic forces of the UAV expressed in the body frame and FBM is the vector of the forces generated by the motors expressed in the body frame. We de nedm as the mass of the UAV, and the vector V B = [u; v; w] T represents the translational speed of the center of mass of the UAV in the body frame and V I = [ _x; _y; _z] T is the translational speed of the center of mass of the UAV in the inertial frame. The angular speed of the UAV with respect to the body frame is de ned as !B = [p; q; r] T . The vector of the gravity force with respect to the inertial frame I is de ned as GI = [0; 0; g] T and with respecrt to the body frame GB = R1B!I 1 GI . For the rotational dynamics we de ne a general vector of the couples in the body frame B = BA + B M + BD + B G , where A is the vector of the couples generated by the aerodynamic forces, M is the vector of the couples generated by the motors, D is the vector of the couples of drag (D) which are opposite to the movement of the propellers and G is the vector of the gyroscopic couples. We de ned the rotational matrix RI!B 1 from the inertial frame (I) to the body frame (B). Denoting roll, pitch, yaw angular displacements as ( ; ; ), we have the rotational matrix RI!B 1 = 0 @ c c c s + c s s s s c c s c s c c s s s c s + c s s s c s c c 1 A (1) where sa = sin(a) y ca = cos(a). To transform the aerodynamic forces of the wind frame to the body frame we have the following rotational matrix RS!B 4 = 0 @ cBcC cBsC sB sC cC 0 cCsB sBsC cB 1 A (2) Where C and B are the angle of sideslip and the angle of attack of the UAV. To transform the forces of the front motors to the body frame we will use the matrix: RM!B 5 = 0 @ c 0 s 0 1 0 s 0 c 1 A (3) Where is the angle of relative movement of the front motors to the body frame B. To transform the aerodynamic forces of the tail towards the body frame we will need the matrix: RS!B 6 = 0 @ cC sC 0 sC cC 0 0 0 1 1 A (4) In order to simplify the mathematical processure we consider that the body of the UAV is symmetrical respect of the axes xB ; yB ; zB and therefore all the inertial products are zero and we will have a diagonal inertia matrix: I = 0 @ Ix 0 0 0 Iy 0 0 0 Iz 1 A (5)"
        ]
    },
    {
        "image_filename": "designv11_100_0000764_1-4020-5123-9_3-Figure3-18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000764_1-4020-5123-9_3-Figure3-18-1.png",
        "caption": "Figure 3-18. Schematic drawing of the geometric model.",
        "texts": [
            " For every simulation step, the energy minimum for the droplet surface is calculated, with the constraint that the centroid of the droplet is fixed at a given location. The surface energy is evaluated and is plotted versus the centroid position. 78 We compare the results to a geometric model, for which the following assumptions have been made: The liquid-air interface does not contribute to the energy change, i.e., its area is approximately constant The base radius of the contact line does not change The contact line always forms a circle (see Fig. 3-18) The potential energy change can then be calculated by evaluating \u2206E(xc) = \u222b rB \u2212rB 2 \u221a r2 B \u2212 \u03be2 \u03b3 (\u03be + xc) d\u03be, (3.27) where rB is the radius of the contact line and xc is the position of the center of the contact line. The radius of the droplet base for a contact angle \u03b8 can be calculated with rB = sin \u03b8 \u00b7 3 \u221a 3V \u03c0(1 \u2212 cos \u03b8)2(2 + cos \u03b8) . (3.28) 79 Chapter 3 The results of the Surface Evolver model are shown in Fig. 3-19 and 3-21. White circles indicate where the contact line arrives at the interdigital edge structure and where it arrives on the bulk pad"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001951_9781118230275.ch8-Figure8.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001951_9781118230275.ch8-Figure8.8-1.png",
        "caption": "FIGURE 8.8 (a) Photograph of the dipstick-type microelectrode array (MEA) consisting of 32 gold microelectrodes on Si (bottom part of the stick) that is connected to a standard sliding contact on a printed circuit board (macroscopic gold pads on the upper left part of the stick; an identical number of contacts is at the back of the stick) via wire bonding (mounted under the black epoxy glue, next to the Au-on-Si-MEA). (b) An enlargement of the Au-on-Si-MEA, when it is immersed in a standard 100-\u03bcl PCR tube. (c) The dipstick connected to the potentiostat. Source: With permission from Reference 91.",
        "texts": [
            " In a recent paper (91), an electrical detection technology has been introduced, which is easy to handle, easy to integrate into automated diagnosis systems, and can also be applied in a conventional manner, for example, as a sensor to detect PCR amplicons directly in the PCR vessel. The work deals with an electrical biosensor detection principle for applications in medical and clinical diagnosis based on the electrically detected displacement assay (EDDA). The electrodes used in the work are shaped as a \u201cdipstick\u201d and can be directly placed into PCR tubes containing the DNA amplicons (Fig. 8.8). The microelectrode array consists of 32 gold electrodes onto which different DNA probes can be immobilized. The detection principle is based on a first 220 BIOSENSORS FOR CLINICAL BIOMARKERS step of hybridization between the immobilized capture probe and a signaling probe labeled with an electroactive moiety (ferrocene). The signaling probe is hybridized to the capture probe at the end carrying the label, whereas at the other end, it presents an extra sequence that is free for the hybridization with the target PCR amplicon"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003081_gt2012-69967-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003081_gt2012-69967-Figure8-1.png",
        "caption": "Figure 8. Torsional deformation of the shaft",
        "texts": [
            " Although this method requires a solid model of a webbed shaft, parameter-based modeling can be used to quickly develop various configurations from a single template model. Geometric elements selected as parameters that can be modified would consist of the base shaft diameter, web thickness and height, and the number of webs. In addition to the torsional stiffness calculation, a solid model also allows for the verification of the motor core weight and the inertia. Fig. 7 provides the meshed FEA model used to determine the rotational deflection used in the stiffness calculation. Figure 8 shows the torsional deflection of the shaft when it is subject to the face constraint at one end and the torque at the other. Figure 9 shows a close-up of the face undergoing maximum deformation. Four points are selected along the periphery of the diameter to obtain an average rotational deformation. The angular deformation is then determined from Eq. (14). l r   (14) It was observed that the torsional stiffness is generally higher as the number of spider bars as well as individual spider bar thickness increases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure20-1.png",
        "caption": "Figure 20 Manifold subsystem (see online version for colours)",
        "texts": [
            " Understand the build volume limitations of the AM systems. Conduct a search for part candidate requirements that have the following characteristics: low production volume, comparative performance requirements, highly complex area and/or difficult to assemble. Once a part candidate is found using weight reduction as a primary driver for redesign. By using weight reduction as a driver, part count decreases, cost will decrease and force the design to become more function based, as opposed to manufacturing based. Figure 20 highlights a complex manifold system for a high performance race car. This part is a good candidate for AM as the part production forecast is only 15 units per year, the manifold simply transfers and reduces low pressure air of ambient air temperature among the different systems within the race car. Due to the encapsulated geometry of the manifold subsystem, several pieces must be constructed to develop the manifold subassembly. The respective bill of materials is highlighted in Table 4. The costs listed were obtained through anonymous suppliers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002308_icems.2013.6754509-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002308_icems.2013.6754509-Figure1-1.png",
        "caption": "Fig. 1 FEM model",
        "texts": [
            " But, proposed model does not consider the influence of leakage flux which flow from one phase to other phases. In this paper, we studied the influence of the leakage flux by using FEM analysis. And we propose a model in consideration of the leakage flux to other phases. And, in order to show validity of the proposed model, we compare simulation results of the proposed model and FEM results. U. COUPLED ANALYSIS OF FEM AND CIRCUIT SIMULATOR In order to explore the influence of leakage flux to other phases, we use general coupled analysis of FEM and circuit simulator. Fig. 1 shows FEM model which consists of central shaft, rotor core, stator core, winding coil, and an air gap. The structure of the SRG is four phase machine where stator and rotor have eight poles and six poles respecting. Fig. 2 shows magnetic characteristic used in FEM model. Fig. 3(a) shows FEM analysis results where A-phase is excited. Fig. 3(b) shows circuit simulation results (A, B and D-phase voltage). We can see from Fig. 3(a), that magnetic flux leaks out to B phase, and Fig. 3(b), when A-phase is excited, B-phase voltage waveform has pulsation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.6-1.png",
        "caption": "Figure 4.6 Design of welded two-step reducer",
        "texts": [
            " 1 0 ) d 2 The three-dimensional design of a casted housing of a one-step reducer with cylindrical gears is presented in Figure 4.5. Welded Housings Awelded layout is applied in a single or low-series production. In principle, the thickness of the walls can be less than those of cast iron housings and covers. It usually takes 30% less than the thickness of cast housings and covers. Housing and housing cover are welded from elements prepared from a sheet. The layout of a welded housing of a coaxial two-stage reducer with welded cylindrical gears is presented in Figure 4.6. During operation, oil in the housing is heated and evaporates. The pressure in the housing grows, leading to disruption of the seals, that is, oil leakage between the seal and the shaft and on the assembly plane, between housing and cover. Considerable heating arises, leading to an increase in lubricant temperature and intensive evaporation. Due to oil splashing caused by the rotation of gears, the cabinet of the housing is full of oil droplets which partially evaporate. Devices for ventilation should prevent both an increase in pressure within the housing and the exhaust of air through the hole, in the form of a vent provided for it"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002780_s10851-011-0286-y-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002780_s10851-011-0286-y-Figure8-1.png",
        "caption": "Fig. 8 Coordinate reference systems Ig, Ib, Ic on the Aibo robot",
        "texts": [
            " |\u03c0(g)| < tol}, where tol is a tolerance limit for the distance between the ground plane and a leg tip point in order to accept it as a support point. 4.2 Coordinate Reference Systems In order to propagate the correction computed to minimize the visual feature error, it is necessary to define the relevant coordinate reference systems coexisting in the robot, and the transformations between them. We denote Ia a generic reference system, and aIb the transformation from Ib to Ia . The three coordinate reference systems of interest in our application, illustrated in Fig. 8 for the Aibo robot, are: \u2013 The fixed reference system whose origin lies on the ground, Ig . Previous works in the literature [19] assume a predefined world reference system. To avoid this limitation, we define this system relative to the basic ground support points defined in Sect. 4.1. \u2013 The body coordinate reference system, Ib, whose origin is the geometrical center of the robot\u2019s body. It is assumed that all the readings from the system configuration (i.e. leg configurations) are provided in this frame of reference"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003073_j.egypro.2012.01.277-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003073_j.egypro.2012.01.277-Figure1-1.png",
        "caption": "Figure 1: Biped heel-and-toe robot kinematics model",
        "texts": [
            " Kinematics modeling includes forward kinematics model and inverse kinematics model, forward kinematics model is to output posture of robot legs in the reference coordinate system according to the given geometric parameters of the various bars and joint movement. Inverse kinematics modeling is output the motions of each joint when the biped heel-and-toe robot\u2019s pose is given. At present, it has too many ways to solve the kinematics problem. In this paper, the kinematics analysis of biped heel-and-toe robot used the more intuitive coordinate transformation method and D-H formula which has been widely used. According to the competition requirements, biped heel-and-toe robot's 6 degrees of freedom are rotational joints. Fig. 1 is the biped heel-and-toe robot kinematics model. In which, i XYZ\u2211 \uff08 \uff09 represent coordinate system, the origins are: right ankle\u2019s projection on the ground , right ankle, right knee, right hip, left hip, left knee, left ankle and left ankle\u2019s 0,1, ,7i = \u22c5\u22c5\u22c5\u22c5 \u22c5 Qiushi Zhang et al. / Energy Procedia 16 (2012) 1799 \u2013 1805 1801 Author name / Energy Procedia 00 (2011) 000\u2013000 jL \uff08 \uff09represent the length of the bar ;1, 2, ,7j = \u22c5\u22c5 \u22c5 \u22c5 \u22c5projection on the ground; j k\u03b8 \uff08 \uff09 represent the angle of the joint "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002406_2011-01-1689-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002406_2011-01-1689-Figure7-1.png",
        "caption": "Figure 7. Numbering of beams for analysis of results",
        "texts": [
            " In order to preserve symmetry an additional elastic wedge is added to the structure to preserve symmetry figure 6. Thus, avoiding important variation of the resonant frequencies of the structure. Frequency response of the base frame with double elastic wedge is compared to the results of the frame without elastic wedges. Multiple sinusoidal input forces of 1N are used in this case. The location of the input forces is indicated in figure 6. Each beam in the structure is numbered as indicated in figure 7. Figure 8 shows the results of the analysis of the structure with and without elastic wedge are compared for each of those 10 beams. The results in figure 8 show that the effect of the elastic wedge depends on the beam considered. For example, the response of beam 1 is not damped but increased. On the other hand, in beams 2 to 5 damping is achieved and the response s decreased. Also, from beams 6 to 10 there is not variation of the response. It was also found in this research that the thinner and longer elastic wedge does not necessarily result in larger damping levels"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003009_imece2013-62977-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003009_imece2013-62977-Figure11-1.png",
        "caption": "Fig. 11: Toy catapult and game [8]",
        "texts": [
            " The research also promoted the development of initial design concepts (i.e. design ideas for how the final product will be configured, function, perform, etc). Part of the research included benchmarking (evaluating similar products) a patent search. A patent search is conducted in order to ensure that no patent is infringed upon as well as to provide the MAE design team ideas as to how they may design their tabletop catapult. The team reviewed numerous patents such as the one represented by the image in fig. 11. The catapult is US patent 6102405. As with each patent reviewed the MAE design team would evaluate the advantages and disadvantages. The MAE design team deemed the advantages of the toy catapult and game to be its simplistic design and the fact that it is compact and portable [9]. Some disadvantages of the patented catapult possessed limited potential energy, no locking mechanism, and the potential failure of the elastic band. Additional research included the administering of a market survey to the WMS students"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003145_iecon.2012.6388702-FigureI-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003145_iecon.2012.6388702-FigureI-1.png",
        "caption": "Figure I. The basic mechanical concept",
        "texts": [
            "9, where data from gyroscope (stabilization system) are depicted (the data are not synchronized with the previous figures). Stabilization error of 0.6 mrad is acceptable for the customer and further improvements have not to be performed. The figure shows another interesting feature in the form of nonsymmetrical friction effect which is caused by imperfections in mechanical components. Fig.10 brings the q - axis current waveforms recorded during stabilization. Following error is minimized thanks to nonlinearities and sampling compensation. (a) (b) Figure I O. Current in q-axis - (a) without and (b) with compensator V. CONCLUSION Hostile and unpredictable - this is the basic characteristic of environment, where described pan/tilt device is doomed to work. Besides construction issues the design of control system when ambitious values of accuracy and repeatability must be met is highly challenging. Advanced techniques based on unique filtering and smart observer for elimination of flexible coupling, backlash and other constraints have been developed and presented in this paper"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.13-1.png",
        "caption": "Fig. 3.13 Double string pendulum",
        "texts": [
            " A dual rod pendulum consists of two rigid massless rods, length l , suspended from frictionless pivots connected to a horizontal rigid massless rod with a point mass, m, at its centre. In the rest position the rigid massless rods, length l , are vertical, as shown in Fig. 3.12a. The motion of the point mass, m, is identical to that of a simple rod pendulum (Sect. 2.3) except that, because of interference between the links, the pendulum angle, , (Fig. 3.12b) cannot exceed 90\u0131. For small amplitudes the time of swing, T , is given approximately by Eq. 2.13 where g is the acceleration due to gravity. A double string pendulum (Fig. 3.13) consists of two simple string pendulums (Fig. 1.4b) arranged in series. The upper inextensible massless string, length l1, is clamped at its upper end, and has a point mass, m1, at its lower end. The lower inextensible massless string, length l2, is attached at its upper end to the point mass m1, and has a point mass, m2, at its lower end. A double string pendulum has four degrees of freedom so possible motions are extremely complicated, and it can display chaotic behaviour. If the point masses m1 and m2 remain in a vertical plane, the planar modes of oscillation are analogous to the modes of oscillation of a double rod pendulum (Sect"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000982_9780471740360.ebs1320-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000982_9780471740360.ebs1320-Figure6-1.png",
        "caption": "Figure 6. SMARTWheel instrumented wheelchair wheel.",
        "texts": [
            " Load cells work the same way but, instead of mounting the strain gages or piezoelectric crystals to a flat plate, they are mounted inside a cylindrical structure that can be secured to objects or surfaces the individuals interact with. Custom force and moment sensing systems using similar technology have also been developed. For instance, wheelchair wheels have been instrumented with strain gages mounted to beams within the plane of the wheel to record the amount of force and torque exerted on the pushrims during wheelchair propulsion (Fig. 6) (20,21). As with kinematic data, the forces and torques recorded from force plates and load cells are susceptible to noise and usually need to be filtered. Low-pass filters are generally used and the cut-off frequencies are determined through a power spectral analysis of the force signals or similar procedure. 2.2.3. Interpretation and Application. Newton\u2019s laws of motion assist with solving problems where forces are of interest to the rehabilitation professional. Suppose you were asked to consult on a case where traction was being prescribed to mobilize a fractured limb"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001576_978-3-642-34381-0_14-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001576_978-3-642-34381-0_14-Figure1-1.png",
        "caption": "Fig. 1. Body frame and inertial frame",
        "texts": [
            " The small-scale helicopter can be described by a hybrid model which contains the nonlinear rigid body dynamics and the dynamics of the rotor. The position of the gravity center of the helicopter, in the inertial frame, is notated by P=[x,y,z]T; the linear velocity of the gravity center in the inertial frame is given byv=[vx,vy,vz]T; the angular velocity of the helicopter represented in the body frame is \u03c9=[p,q,r]T; the attitude is denoted by Euler angles \u03b7=[\u03c6,\u03b8,\u03c8]T. The body frame and inertial frame are shown in Fig.1. Actually, using the Euler angles represent the attitude kinematics, there are singularities when the pitch angle equal to\u00b190\u00b0. Here, it is assumed that the flight condition will never reach the singularity. The nonlinear rigid body dynamics in terms of the translational and rotational dynamics are given by: ( )m mg I I  =  = +    = \u2212 \u00d7 + b 3 P v v R \u03b7 f e \u03b7 = \u03c0\u03c9 \u03c9 \u03c9 \u03c9 M     (1) Where the mis the mass of the body and I is the inertia matrix.e3is a unit vector along the zaxis of the inertia frame"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.9-1.png",
        "caption": "Fig. 2.9 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRPRR (a) and 4RRRPR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R||R||P||R\\R (a) and R||R||R||P\\R (b)",
        "texts": [
            "5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10. 4PRRRR (Fig. 2.6b) P||R\\R||R\\R (Fig. 2.1i) Idem No. 9 11. 4RRPRR (Fig. 2.7a) R\\R\\P\\kR\\R (Fig. 2.1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2.8a) P||R||R||R||R\\R (Fig. 2.1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig. 2.9a) R||R||P||R\\R (Fig. 2.1n) Idem No. 13 16. 4RRRPR (Fig. 2.9b) R||R||R||P\\R (Fig. 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No. 13 18. 4RRPRR (Fig. 2.10b) R||R||P||R\\R (Fig. 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2.1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.54-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.54-1.png",
        "caption": "Fig. 2.54 4PaPaPR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa||Pa||P\\R",
        "texts": [
            " 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig. 2.49) R\\Pa||Pa\\kR (Fig. 2.22j) The first and the last revolute joints of the four limbs are parallel 38. 4RPaPaR (Fig. 2.50) R\\Pa||Pa\\kR (Fig. 2.22k) Idem No. 37 39. 4RPaRPa (Fig. 2.51) R\\Pa\\kR\\Pa (Fig. 2.22l) Idem No. 15 40. 4PaRPPa (Fig. 2.52) Pa\\R\\P||Pa (Fig. 2.22m) Idem No. 34 41. 4PaPaRP (Fig. 2.53) Pa||Pa\\R\\P (Fig. 2.22n) The axes of the revolute joints connecting links 7 and 8 of the four limbs are parallel 42. 4PaPaPR (Fig. 2.54) Pa||Pa||P\\R (Fig. 2.22o) Idem No. 4 43. 4PRPaPa (Fig. 2.55) P\\R\\Pa||Pa (Fig. 2.22p) Idem No. 5 44. 4PPaRPa (Fig. 2.56) P||Pa\\R\\Pa (Fig. 2.22q) Idem No. 33 45. 4PPaPaR (Fig. 2.57) P||Pa||Pa\\R (Fig. 2.22r) Idem No. 4 46. 4PaPPaR (Fig. 2.58) Pa||P||Pa\\R (Fig. 2.22s) Idem No. 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002436_afrcon.2013.6757663-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002436_afrcon.2013.6757663-Figure2-1.png",
        "caption": "Fig. 2. Self collision avoidance parameters",
        "texts": [
            " For arm segments with a collision distance d, self collision avoidance is implemented by adjusting the glenohumeral abduction and /or glenohumeral rotation joint angles, of the humanoids. First, collision avoidance for the upper arm is obtained, followed by collision avoidance for the lower arm. For the legs, the hip abduction and /or knee flexion joint angles are adjusted. For either the arm or the leg, collision avoidance is attained, first, by adjusting the relevant abduction joint angle \u03b8U by the collision avoidance angle \u03b8col: \u03b8U = \u03b8U \u2212 \u03b8col \u00d7 (|d|+ 1) dS (13) where the distance dS ,(shown in Figure 2), of the collision point from the glenohumeral joint or the hip joint is: dS = \u221a pc \u2212 pS l (14) where pS is the position of the joint and l is the length of the segment. Using the above equations, the collision avoidance angle \u03b8col is first scaled by the magnitude of the collision distance to allow the colliding segments to move away from each other. \u03b8col is then scaled by the inverse of dS , the distance of the collision point from the joint, to ensure that the resultant posture remains close to the original posture"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002439_amm.378.329-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002439_amm.378.329-Figure1-1.png",
        "caption": "Fig. 1 Coordinate system of PCE clutch when forced continuous action",
        "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. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-21/05/15,20:32:48) movement. They are the translation around the X, Y axis and the rotation around Z axis. To facilitate analysis, it isn\u2019t considering circumferential displacement of the sprag. Therefore, the movement of sprag can be simplified as the radial displacement and rotation around the own center. The coordinate system of PCE clutch forced continuous action is shown in figure 1.The base coordinate system is X1O1Y1; the design coordinate system of sprag is X2O2Y2. The specific form of the PCE clutch forced continuous phenomenon is the contact between sprag flange and the after shoulder of the adjacent sprag. In Fig 1, point K is the circle arc of the sprag flange. In coordinate system X2O2Y2, the coordinates of point K are 2 2O O T K K[x ,y ,0 ,1 ] and the radius is rk. M \u2032 , N \u2032 are the two endpoints on the straight line segment of the adjacent sprag after shoulder. Suppose d is the linear distance between point K and lineM N\u2032 \u2032 . So, the sufficient and necessary conditions to judge the contact of the adjacent sprag are shown as below. (1) Kd r= ; (2) Projection point K \u2032 of K point to the line M N\u2032 \u2032 is inside the segment M N\u2032 \u2032 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003535_eucap.2012.6205942-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003535_eucap.2012.6205942-Figure2-1.png",
        "caption": "Fig. 2 (a) Leaky wave structure based on two spiraphase-type reflectarrays, (b) Reactively-loaded spiraphase-type element",
        "texts": [
            " According to (5), the investigated LWS would demonstrate a conical radiation pattern with aperture angle 02 . The same conclusion is stated in [3,8]. Thus, it is possible to obtain reconfigurable conical pattern with two phase agile surfaces separated by a distance h. In this work spiraphase-type reflectarrays are used to control phases t and b of the reflection coefficients. III. GEOMETRY OF THE LWS The investigated reconfigurable LWS based on two spiraphase-type reflectarrays separated by a distance h is shown in Fig. 2 (a). Both reflectarrays contain a frequency selective surface (FSS) and a metal screen situated at a distance hs from the FSS. Each FSS is formed by reactively loaded ring slot spiraphase elements (similar to described in [9]) arranged at the nodes of square grid with period b. Circularly-polarized feed is integrated to the bottom FSS. The top reflectarray is partially transparent due to the small ring slot apertures of inner radius ra and outer radius rb, perforated in the metal screen. The reactively loaded ring slot spiraphase element is shown in Fig. 2 (b). It contains loaded ring slot with inner radius r1 and outer radius r2. Two reactive loads are connected to the ring slot at the angular position The geometry of the element is optimized in a such way that for the normally incident linearly polarized wave with polarization plane parallel to the axis AA1 the FSS is equivalent to the capacitance with normalized reactance of -1. For the linearly-polarized wave with polarization plane orthogonal to axis AA1 this FSS is equivalent to the inductance with normalized reactance of 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002125_978-3-642-23147-6_34-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002125_978-3-642-23147-6_34-Figure1-1.png",
        "caption": "Fig. 1. The velocity control of the robots\u2019 movements",
        "texts": [
            " Another research which concentrates on mobile robot path planning is described using the potential field method to speed up the mobile robot planning. This approach was still able to track a moving object even though there were moving obstacles in surroundings [7]. They kept the potential function in its traditional way, and derived an explicit function for controlling the robot speed at a certain stage such as the speed of the target (obstacle) and relative positions between the obstacles, the goal and between the robots themselves. The speed of the robot function can be determined as following (refer to Fig. 1). cos \u2211 \u03b2 cos sin (1) Where: = velocity of the robots = velocity of the object tar = velocity for the robots to track the object = angle of tar = angle of = angle of the relative position from the robot to the target = scaling factor = the relative motion between the robot and the target and are the speed and direction of ith obstacle Reference [8] examines the realization and visualization of ball passing, and moveto-ball behavior of a robotic soccer game. This research describes three identical mechanical mobile robots with a formation ready to pass a ball cyclically in a zigzag pattern"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001531_s00707-013-0995-y-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001531_s00707-013-0995-y-Figure4-1.png",
        "caption": "Fig. 4 Two contiguous rigid beams",
        "texts": [],
        "surrounding_texts": [
            "conditions by successively degenerating the derivation grade, once more changing sign for odd derivatives. (This result reflects the procedure of integrating by parts, see also Eq. (33), with its consecutive sign changing). For the present example, the procedure is demonstrated in Table 1: the first row indicates the initial terms, i.e., D and dQc = [dQc/dx]dx , and the first column shows the resulting operators that are applied to [dQc/dx]. The result is displayed in the second column.\n2.1.1 Summary procedure\nThe foregoing is not restricted to only one spatial variable nor to one only deflection function (e.g., a plate w(x, y, t): one function, two spatial variables, or a beam qT = [\u03d1(x, t), v(x, t), w(x, t)]: one torsional and two bending functions, one spatial variable, see [7]). The latter case is summarized in Table 2.\n2.2 Elastic deflection suppressed\nWhen disregarding elastic deflections, we are left with\ny\u0307 = ( vox voy \u03c9oz )T (12)",
            "which characterizes the plane motion of a rigid body. The corresponding virtual work then reads\u222b B \u03b4yT[dM11y\u0308 + dG11y\u0307 \u2212 dQ1] = 0 with Eq. (6) et seq.\n\u222b B \u03b4yT \u23a7\u23a8 \u23a9 \u23a1 \u23a3 dm 0 \u2212dm v 0 dm dm x \u2212dm v dm x dJ o \u23a4 \u23a6 y\u0308+ \u23a1 \u23a3 0 \u2212dm\u03c9oz \u2212dm x\u03c9oz dm\u03c9oz 0 \u2212dm v\u03c9oz dm x\u03c9oz dm v\u03c9oz 0 \u23a4 \u23a6 y\u0307\u2212dQ1 \u23ab\u23ac \u23ad=0. (13)\nSince v = 0, spatial integration yields\n\u03b4yT \u23a7\u23a8 \u23a9 \u23a1 \u23a3m 0 0\n0 m m c 0 m c J o\n\u23a4 \u23a6 y\u0308 + \u23a1 \u23a3 0 \u2212m\u03c9oz \u2212mc\u03c9oz\nm\u03c9oz 0 0 mc \u03c9oz 0 0\n\u23a4 \u23a6 y\u0307 \u2212 Q \u23ab\u23ac \u23ad = 0\n:= \u03b4yT[My\u0308 + Gy\u0307 \u2212 Q] (14)\nwhere c: mass center distance.\n3 Multi-beam systems\n3.1 Rigid beam systems\nEquation (14) is now used to calculate a system of rigid bodies. This is easily done by summing up their individual virtual work according to Eq. (14):\nN\u2211 k=1 \u03b4yT k [ M y\u0308 + G y\u0307 \u2212Q ] k = 0 = \u03b4qT N\u2211 k=1\n( \u2202 y\u0307k\n\u2202q\u0307\n)T [ My\u0308 + Gy\u0307 \u2212 Q ] k . (15)\nThe mass center velocities of frame origin (i) are those of frame (p) at the coupling point L (p: predecessor of i), which are to be transformed into the actual frame and superimposed with the relative velocities of frame (i).",
            "and henceforth called the \u201ckinematic chain.\u201d Starting with i = 1, which does not have a predecessor, one obtains\n\u239b \u239c\u239c\u239d y\u03071 y\u03072 ...\ny\u0307N\n\u239e \u239f\u239f\u23a0 = \u23a1 \u23a2\u23a2\u23a3 E T21 E ... ... . . .\nTN1 TN2 \u00b7 \u00b7 \u00b7 E\n\u23a4 \u23a5\u23a5\u23a6 \u239b \u239c\u239c\u239d y\u03071r y\u03072r ...\ny\u0307Nr\n\u239e \u239f\u239f\u23a0 (18)\nwhere Ti j = Ti p(i) \u00d7 \u00b7 \u00b7 \u00b7 Ts( j) j (p: predecessor, s: successor). Next, the relative velocities are computed with y\u0307i,rel = Fi q\u0307i (Fi : local functional matrices) yielding\n\u239b \u239c\u239c\u239d y\u03071 y\u03072 ...\ny\u0307N\n\u239e \u239f\u239f\u23a0 = \u23a1 \u23a2\u23a2\u23a3 F1 T21F1 F2 ... ... . . .\nTN1F1 TN2F2 \u00b7 \u00b7 \u00b7 FN\n\u23a4 \u23a5\u23a5\u23a6 \u239b \u239c\u239c\u239d q\u03071 q\u03072 ...\nq\u0307N\n\u239e \u239f\u239f\u23a0 . (19)\nEquation (15) reads, in vector notation,\n[( \u2202 y\u03071\n\u2202q\u0307\n)T ( \u2202 y\u03072\n\u2202q\u0307\n)T \u00b7 \u00b7 \u00b7 ( \u2202 y\u0307N\n\u2202q\u0307 )T ]\u23a1\u23a2\u23a2\u23a2\u23a3 M1y\u03081 + G1y\u03071 \u2212 Q1 M2y\u03082 + G2y\u03072 \u2212 Q2 ...\nMN y\u0308N + GN y\u0307N \u2212 QN\n\u23a4 \u23a5\u23a5\u23a5\u23a6 = 0 (20)\nwhere q\u0307T := (q\u0307T 1 , q\u0307T 2 . . .) is defined with Eq. (19).\n3.1.1 Summary procedure\nInserting Eq. (19) into Eq. (20) yields a \u201cGaussian form\u201d according to Table 3. All non-marked elements are zero. Disposing thus of a representation with an upper (block-) triangular matrix enables to solve the motion equations recursively without inversion of the (global) mass matrix."
        ]
    },
    {
        "image_filename": "designv11_100_0002756_scored.2012.6518622-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002756_scored.2012.6518622-Figure1-1.png",
        "caption": "Figure 1. TRMS Diagram",
        "texts": [
            " For this reason, the tail which mounted normal to yaw plane operates to counter the effect. This behavior recognized as coupling effect and giving the challenging to engineer and researcher in controlling the characteristic. Counter balance mass is equipped as stabilizer and replicate the helicopter cockpit. It hanging to cylinder beam and mounted same as center of gyroscopic movement. Twin rotor defines as multi input and multi output as it consists two inputs which come from two DC motors. Two outputs are identified as pitch and yaw angle. The schematic diagram exhibits as Fig 1. Dynamic equation of twin rotor system is derived based on Newton Law. The inputs of the system for pitch planes are 1\u03c4 , M1, MFG, Mb\u03c1, and MG, which are the torque induce, the nonlinear static characteristic of motor, the gravity momentum, the friction forces momentum and the gyroscopic momentum respectively. And the utout angle displacement of pitch is \u03c1 and the dynamic equation of twin rotor as follows: )( 11 uf=\u03c4 (1) )( 11 \u03c4fM = (2) ),,,,,( \u03b3\u03b3\u03b3\u03c1\u03c1\u03c1fM FG = (3) ),,,,,( \u03b3\u03b3\u03b3\u03c1\u03c1\u03c1\u03c1 fM b = (4) ),,,,,( \u03b3\u03b3\u03b3\u03c1\u03c1\u03c1fM G = (5) Where f is a function of \u03c1 and \u03b3 , their derivatives and 1u is pitch control signal"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003785_icnmm2013-73187-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003785_icnmm2013-73187-Figure4-1.png",
        "caption": "Fig. 4 Microbeads sorting apparatus",
        "texts": [
            " Finally, the droplet is torn by a combined effect of its own weight and a meniscus force of the liquid surface. When the flow rate is increased, the last liquid droplet ejected sticks to the previous droplet before forming the microbead. As the liquid surface reaches the nozzle, a row of microbeads is continuously produced. At high flow rates, the microbeads formed are bigger in size and the shape resembles pendant droplets. To select a certain microbead, we observed the particle diameter and bead shape on a sorting plate using the USB video camera with a zoom lens attached as shown in Fig. 4. We separated and selected the microbeads individually using the sorting plate. 2 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/26/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use This research has led to a new method for the stable production of uniform microbeads, however, most of the microbeads produced by the new method were not perfectly spherical but more like a pendant droplet. At high flow rates, the pendant drops became larger and the microbeads had a pointed tip"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002704_tai.1958.6367332-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002704_tai.1958.6367332-Figure1-1.png",
        "caption": "Fig. 1. Stator and rotor windings",
        "texts": [
            " x = displacement transfer func\u00ae(s) tion given by equation 23 F=the frictional torque per unit angular velocity /iW\u00bb M)\u2014 functions of time denned by equations 14 and 15 g(t) \u2014 modulating function in control phase G($) = the Laplace transform of gif) H (s) =\u00bb \u039b = velocity transfer function given by equation 22 7=the rms value of the current in the reference phase *i\u00bb *2 = instantaneous values of stator currents ia, ih \u2014 instantaneous values of rotor currents Ha, isb = rotor currents giving same mag- netomotive force as stator currents / = t h e moment of inertia of the rotor K \u2014 torque constant introduced in equation 5 L = inductance of a rotor winding M = maximum value of mutual inductance between a stator and a rotor winding p = d/dt r = the resistance of a rotor winding Te = electric torque Tm = mechanical torque \u0398 = instantaneous value of angle between rotor and stator winding \u03b8' =\u03c1\u03b8 \u03b8\" =\u03c12\u03b8 \u00ae(s)=the Laplace transform of \u0398 \u00ae'(s)= the Laplace transform of \u0398' re = L/r is the rotor electrical time constant Tm = J/F is the rotor mechanical time constant \u03c9 = 2\u03c7/, where / is the frequency of the reference current Fundamental Equations With reference to Fig. 1, let 1 and 2 be the stator windings, a and b the rotor windings, respectively; thus, ia, \u00f9 are the instantaneous values of the rotor currents, and ih 12 are the instantaneous values of the stator currents. If tsa and isi, are defined as components of rotor currents wluch give the same magnetomotive force as the stator currents, then, isa \u2014i\\ cos 0+i*2 sin 0 and isb \u2014 iz cos \u0398\u2014ii sin 0 (1) (2) Since the rotor windings, are closed, the differential equations for the rotor windings are given by 0 = (r+Lp)ia+Mpi,a and 0 = (r+Lp)ib+Mpisb (3) (4) The electric torque is proportional to the vector product of the currents in the a and b axes and is thus given by Te = KM2 [iai$b\u2014hisa ] (5) From equations 3, 4, and 5, the electric torque is given by KM* r ",
            " Statement of the Problem The air-turbine drive which was to form the application for the new control system was in existence a t the t ime the program was initiated. The drive control system a t t h a t t ime comprised a mechanical governor and a hydraulic actuat ion system for controlling the position of a throt t le valve which, in turn , controlled engine compressor bleed air flow to the turbine. T h e open-loop block diagram of the drive itself, wi thout controls, is essentially as shown in Fig. 1. T h e hydraulic ac tuator and control valve for controlling air flow to the turbine is represented as an integration having a gain Kh t he Laplace transform notat ion being used. The inp u t to this block is control oil flow and the ou tpu t is in the form of an aerodynamic torque on the turbine. The gain K\\ is. subject to relatively wide variat ions in magni tude depending upon the enginecompressor bleed energy condition avai l - able a t the turbine. An unusua l feature of the speed-controli sys tem of an air-turbine drive using jetengine compressor bleed air is t h a t it isreally m a n y systems. Wi th reference to Fig. 1, the gain K\\ of the integration between control oil flow and turbine torqueis a variable, depending primari ly uponcompressor discharge pressure, turbineexhaust pressure (airplane al t i tude), and^ A n Electro-Hydraulic Speed and Load Division Control for Constant-Speed Air-Turbine Drives MAY 1958 Dantowitz, Norris\u2014Speed and Load Control for Air-Turbine Drives 99."
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.5-1.png",
        "caption": "Fig. 3.5",
        "texts": [
            "5): aP = dvP dt = \u03c9\u0307 \u00d7 rAP + \u03c9 \u00d7 r\u0307AP . Since point A is fixed in space (r\u0307A \u2261 0), we have r\u0307AP = r\u0307P = 3.1 Kinematics 133 = vP = \u03c9 \u00d7 rAP . Thus, aP = \u03c9\u0307 \u00d7 rAP + \u03c9 \u00d7 (\u03c9 \u00d7 rAP ) . (3.6) Equations (3.5) and (3.6) reduce to (3.2a,b) in the special case of a rotation about a fixed axis. The general motion of a rigid body can be understood as a composition of a rotation and a translation. To show this, we first consider the case of plane motion, where all the particles move in the x, y-plane or in a plane parallel to it (Fig. 3.5a). Position vectors to arbitrary points P and A which are fixed in the body are connected by rP = rA + rAP . Let us introduce the unit vectors er (in the direction from A to P ) and e\u03d5 (perpendicular to rAP ). They are also fixed in the body and therefore move with the body. Since rAP = r er, we can write rP = rA + r er. Note that r = const. Therefore, differentiation yields r\u0307P = r\u0307A + re\u0307r . The vector e\u0307r is found through the following considerations. If the vector rAP rotates through the angle d\u03d5 during the infinitesimal time interval dt, then the vectors er and e\u03d5 are also rotated by d\u03d5. According to Fig. 3.5b we have der = d\u03d5e\u03d5 and therefore e\u0307r = der/dt = \u03d5\u0307e\u03d5. Similarly, we obtain e\u0307\u03d5 = \u2212\u03d5\u0307er (cf. Section 1.1.4). Thus, the velocity of P is given by r\u0307P = r\u0307A + r \u03c9 e\u03d5 where \u03c9 = \u03d5\u0307, and the acceleration is found to be r\u0308P = r\u0308A + r \u03c9\u0307 e\u03d5 + r \u03c9 e\u0307\u03d5 = r\u0308A + r \u03c9\u0307 e\u03d5 \u2212 r\u03c92er . In summary we have rP = rA + rAP , vP = vA + vAP , aP = aA + ar AP + a\u03d5 AP (3.7a) with rAP = r er, vAP = r \u03c9 e\u03d5, ar AP = \u2212 r\u03c92er, a\u03d5 AP = r \u03c9\u0307 e\u03d5 . (3.7b) Each of the relations (3.7a) consists of two parts. The quantities rA, vA and aA represent the translation of the body, whereas the other terms (see (3",
            " The vectors vAP and a\u03d5 AP are perpendicular to rAP ; the vector ar AP points in the direction from P to A (centripetal acceleration). Thus, the velocity (acceleration) of an arbitrary point P is equal to the sum of the velocity (acceleration) 3.1 Kinematics 135 of an arbitrary point A and the velocity (acceleration) of point P due to the rotation about A. Frequently we have to express the velocity and acceleration of P in a Cartesian coordinate system. To do so, we first write down the coordinates of P (see Fig. 3.5a): xP = xA + r cos\u03d5, yP = yA + r sin \u03d5 . If we differentiate once (\u03d5 = \u03d5(t); chain rule!) we obtain the components of the velocity vector and a second time those of the acceleration vector (\u03d5\u0307 = \u03c9): vPx = x\u0307P = x\u0307A \u2212 r\u03c9 sin \u03d5, vPy = y\u0307P = y\u0307A + r\u03c9 cos\u03d5, aPx = x\u0308P = x\u0308A \u2212 r\u03c9\u0307 sin \u03d5\u2212 r\u03c92 cos\u03d5, aPy = y\u0308P = y\u0308A + r\u03c9\u0307 cos\u03d5\u2212 r\u03c92 sin \u03d5 . The meaning of the individual terms can be seen in Figs. 3.5c,d. Equations (3.7a,b) can be used to determine the velocity (acceleration) of point P graphically with the aid of a velocity diagram (acceleration diagram)",
            " The corresponding velocity and acceleration, respectively, of a point P are given by (3.5) and (3.6). In addition, we have to account for the velocity vA and acceleration aA of point A (i.e., of the x\u0304, y\u0304, z\u0304-system) with respect to the fixed system x, y, z. Thus, the general motion of a rigid body in space is described by rP = rA + rAP , vP = vA + \u03c9 \u00d7 rAP , aP = aA + \u03c9\u0307 \u00d7 rAP + \u03c9 \u00d7 (\u03c9 \u00d7 rAP ). (3.8) rP rA x y rAP x\u0304 P z A z\u0304 y\u0304 Fig. 3.8 Equations (3.8) are also valid in the case of a plane motion. Assuming that the motion takes place in the x, y-plane (cf. Fig. 3.5a), we can write \u03c9 = \u03c9 ez, \u03c9\u0307 = \u03c9\u0307 ez, rAP = r er . Inserting into (3.8), we obtain Equations (3.7a,b) for the position, velocity and acceleration of P : rP = rA + r er, vP = vA + \u03c9 ez \u00d7 r er = vA + r\u03c9 e\u03d5, aP = aA + \u03c9\u0307 ez \u00d7 r er + \u03c9 ez \u00d7 (\u03c9 ez \u00d7 r er) = aA + r\u03c9\u0307 e\u03d5 + \u03c9 ez \u00d7 r\u03c9 e\u03d5 = aA + r\u03c9\u0307 e\u03d5 \u2212 r\u03c92er . 3.1 Kinematics 139 E3.1Example 3.1 A slider crank mechanism consists of a crankshaft 0A and a connecting rod AK (Fig. 3.9a). The crankshaft rotates with a constant angular velocity \u03c90. Determine the angular velocity and the angular acceleration of the connecting rod as well as the velocity and the acceleration of the piston K in an arbitrary position"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003067_amr.199-200.824-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003067_amr.199-200.824-Figure1-1.png",
        "caption": "Fig. 1 Torsional model of coupling rotor system",
        "texts": [
            " In this study, a two degree of freedom model is developed to investigate clearance nonlinearity of torsional vibration in a coupling rotor system. The proposed model is applicable for any kind of coupling-rotor mechanisms with clearance such as rotor-bearing, clutch-rotor and gear-rotor systems. The main focus of this research is to investigate the vibration due to coupling effects as well as change of clearance spring stiffness in both symmetric and asymmetric models by simulation and experiment. The physical representation of torsional vibration of a coupling-rotor system with clearance is shown in Fig. 1. Power transmits from engine through clutch, gear or coupling to the rotor. Usually engine provides fluctuating torque to the system which causes torsional vibration in the system. Static and dynamic imbalances due to misalignment or eccentricity have been neglected. So, the physical model (Fig.1) is adapted to a simplified translational two degree of freedom model including vibro-impact (Fig. 2). Usually vibro-impact is modeled through the consideration of coefficient of restitution. In this paper, the rigid part of the driving side coupling is converted to a spring of very high stiffness 3k and damper 3c . Such consideration helps to preserve the effects of restitution as well as the rigidity of the driving side coupling part. 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_100_0003648_icdma.2012.194-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003648_icdma.2012.194-Figure4-1.png",
        "caption": "Figure 4. Direct method",
        "texts": [
            " after established the parameters: = = + 2) Direct method The so-called direct method is to directly generate involute with the involute equation. The specific process is: establish parallel curve 1 and parallel curve 2 respectively in plane yz and plane zx which corresponds to fog x and fog y with the parallel curve function of CATIA. And then, mix parallel curve 1 with parallel curve 2 for one curve through mixture function. Finally, project the mixed curve on plane xy to get the standard involute, as shown in Fig. 4. The involute will always have an endpoint located on (0, ) whether through tracing method or direct method, when the base circle fell on the lateral of dedendum circle, we need to make the endpoint of involute located inside of the base circle by extrapolation, avoid the endpoint located on (0, ). This is the only way to make sure that there are no mistakes when generate the transition fillet because of the involute do not intersect with the dedendum circle, in other words (the spur gear as example): > ( ) When = 20\u00b0 = 1 c = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001892_9781118354162.ch19-Figure19.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001892_9781118354162.ch19-Figure19.16-1.png",
        "caption": "Figure 19.16 Setup of an optical fiber DNA sensor. The sensing component is a piece of optical fiber onto which the capture probe is immobilized. The upper part of the figure shows a fluorescence microscope that includes the excitation source (an Ar laser), the photomultiplier light detector (PM) and a dichroic mirror that resolves excitation and fluorescence beams. Sample solutions, staining dye (ethidium bromide-EB) and hybridization buffer are delivered by syringes to the hybridization cell. Adapted with permission from [47]. Copyright 1994 Elsevier.",
        "texts": [
            " In this stage of the process, the sample analyte displaces a fraction of the labeled analog, leading to a decrease in the fluorescence. Alternatively, transduction in optical nucleic acid sensors can be carried out in the sandwich format. In this approach, the probe is designed so that after hybridization with the target, an overhanging target fragment remains in the single-strand form. This fragment is then used to bind by hybridization a signaling probe tagged with a fluorescent label. As shown in Figure 19.16, an intrinsic fiber optic DNA sensor has been obtained by immobilization of the capture probe onto a piece of decladded 400-mm diameter quartz fiber [47]. To this end, the fiber surface is first activated with a long-chain aliphatic spacer terminated in 50-O-dimethoxytrityl-20-deoxyribonucleoside. Starting with this molecular unit, solid-phase DNA synthesis has been performed at the fiber surface in order to obtain the capture probe. The transduction setup (Figure 19.16) includes the excitation source (an Ar laser) and a photomultiplier light detector. Discrimination of excitation and emission beams is achieved by a dichroic mirror. Hybridization is conducted in a small plastic cell that contains the sample diluted with the hybridization buffer. After hybridization the cell is flushed with the buffer and the staining of the resulting duplex is achieved by incubation with the ethidium bromide intercalating dye. Finally, unreacted dye is flushed away by a stream of fresh buffer and the fluorescence signal is measured"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003525_20120905-3-hr-2030.00102-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003525_20120905-3-hr-2030.00102-Figure2-1.png",
        "caption": "Figure 2: EMPS Components",
        "texts": [
            "1 Experimental setup The EMPS is a high-precision linear Electro-Mechanical Positioning System (see Fig. 1). It is a standard configuration of a drive system for prismatic joint of robots or machine tools. Its main components are: - A Maxon DC motor equipped with an incremental encoder to make a PD motor position control. - A Star high-precision low-friction ball screw drive positioning unit. An incremental encoder measures the angular position of the screw. - A carriage which moves a payload in translation. These components are presented in Fig. 2. All variables and parameters are given in SI units on the load side. 2.2 Inverse dynamic model of the flexible joint robot The mechanical system has 2n = dof, one rigid dof 1q and one flexible dof 2q where: 1q (m/s) is the motor position, 2q (m/s) is the flexible dof position where 12 1 2q q q= + is the load position. The dynamic model does not depend on gravity such that the dynamic parameters are two masses, one spring, and 4 friction coefficients (see Fig. 3). The Inverse Dynamic Model (IDM) calculates the motor force according to the joint positions and their derivatives"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003518_amr.655-657.1023-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003518_amr.655-657.1023-Figure1-1.png",
        "caption": "Fig. 1 Tested robot configuration and its frame",
        "texts": [
            " And literature [8] proposed a two-step identification method by separating the angle errors and length errors to identify the parameter errors respectively in the whole process. This paper discusses a simple and practical calibration method of the robot kinematic parameters, which considers the practicability of the method in robot development and combines the DH model with distance-based error model in order to indentify parameter errors expediently. The basic idea of kinematic parameter identification The kinematics model of robot manipulator. The joint-type 6R robot that studied in this paper is presented in Figure 1. The transformation matrix i-1 Ai between the adjac- ent link can be expressed as follows: i-1 Ai= Rot(x,\u03b1i-1) Trans(x,ai-1) Rot(z,\u03b8i) Trans(z,di) (1) In this way, the matrix 0 T6 which describes position and pose of end-effector in the base frame of robot, can be obtained by continually product by matrix i-1 Ai (the difference between the general expression and the new expression is that 5 parameters were used here ): 6 5 5 4 4 3 3 2 2 1 1 0 6 0 AAAAAA\u03a4 = (2) Formula 2 expresses the robot mapping relationship from the joint space to operating space, it is named as Kinematic equation and from which the end-efector position px\uff0cpy\uff0cpz and orienta -tion angle \u03b1z, \u03b2y, \u03b3x relative to the robot base frame can be obtained by giving the joint angle \u03b8i ",
            " Experimental study of kinematic parameter identification A renishaw laser interferometer ML10 was used to measure a 6R robot end-effector displacement distance. Supposing other axes coordinates value does not change, the end-effector displacement distance was measured along robot Cartesian coordinate system axis direction between 2 points and Fig.2 Experimental equipment configuration was recorded respectively to the measured points, and all recorded data were taken as basis for identifying the robot kinematics parameter errors. Establishing a 6R robot kinematics model (as shown in Figure 1), The test process described as following. Choosing an point in the robot's work space as a reference point, and on this point, the end-effector position and poseture must be easily to install the laser spectroscopy and reflector. Relative to the reference point, the end-effector moves along single axis of Cartesian coordinates and 20 points were taken as measuring point in a movement trace, in the moving process along a trace, the end-effector poseture and the other two terminal axial positions were kept to be invariable"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003302_amr.487.327-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003302_amr.487.327-Figure11-1.png",
        "caption": "Fig. 11 The filling process: STEP=580",
        "texts": [
            " The block in Figure 6 is the virtual mold. After that, the materials, contact conditions, boundary conditions, gravity, initial conditions and the operating parameters are defined respectively to complete the pre-treatment of the numerical simulation of the filling and the solidification process of the air intake hood [4][5][6]. Then, all the defined parameters were saved and the dynamic simulation of the filling and solidification process of the air intake hood could be operated. Figure 7, Figure 8, Figure 9, Figure 10, Figure 11 and Figure 12 are the results of the FEM simulation of the filling and solidification conditions at different moments. It can be seen through the simulation of the casting filling and solidification process of the air intake hood that the alloy liquid was poured into the bigger hole of the air intake hood first. Then the rest began to be filled after the big hole was full (STEP=100), until all parts were full. Because of the contract effect with the plaster model, the temperature of the alloy liquid in the front of the air intake hood (the location of the thin-walled away from the gate) was decreased rapidly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001687_978-3-7091-1187-1_5-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001687_978-3-7091-1187-1_5-Figure11-1.png",
        "caption": "Figure 11. Stresses \u03c313 (x1, x3, t) computed with linear elastic model (a) and with elasto-plastic model (b), when t=0.7547 s, excitation at linear natural frequency with 500 N m\u22121 amplitude.",
        "texts": [
            "827 rad/s, and the amplitude of the distributed transverse force is 500 Nm \u22121. The figure shows the transverse displacement of the middle beam point along two excitation cycles and the projection of the trajectory on the phase plane defined by the transverse displacement and velocity. Te represents the excitation period. In this case, both models predict periodic oscillations dominated by the excitation frequency but the elasto-plastic model predicts smaller displacement and velocity amplitudes than the elastic model. Figure 10 and Figure 11 show the longitudinal and shear stresses, \u03c311 and \u03c313, computed with the linear elastic model and with the elasto-plastic model, both models geometrically nonlinear, at t= 0.7547 s. The difference between values computed with the different models is higher near the clamps and in the middle of the beam. Due to plasticity, the surfaces defined by \u03c311 and \u03c313 become uneven near the clamps (as occurred in the former example, see Figure 8). Although plastic strains also alter the stresses in statics, this unevenness effect was only found in dynamics and is due to the oscillations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.7-1.png",
        "caption": "Figure 4.7 Vent with cap",
        "texts": [
            " Considerable heating arises, leading to an increase in lubricant temperature and intensive evaporation. Due to oil splashing caused by the rotation of gears, the cabinet of the housing is full of oil droplets which partially evaporate. Devices for ventilation should prevent both an increase in pressure within the housing and the exhaust of air through the hole, in the form of a vent provided for it. For units with a larger amount of oil, a special cover is used which prevents oil particles from passing through it (Figure 4.7). Vents consisting of a casing, cap, ring and wire mesh are also used for ventilation. The oil level indicator, as a necessary part of the driver, is, for design reasons, frequently used as a vent. In any system of gear lubrication there is oil in the housing of the driver, thus the housing is used as a reservoir. In drive operation the oil grows old, gets polluted by resins extracted from the oil, and tooth wear products arise too. Thus, oil gradually loses its properties and should be replaced periodically"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003161_amm.105-107.244-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003161_amm.105-107.244-Figure1-1.png",
        "caption": "Figure 1 CAD model of test lifting platform",
        "texts": [
            " The contribution presents such method using application of three different boundary conditions which allow describing the nonlinear behavior of system by tools of linear analysis. Consequently, the modal analysis of the entire system is performed for all three types of boundary conditions. On the base of mode shapes visualization the proper mode shapes are extracted which satisfy the real behavior of the lifting platform. The subject of the contribution is the test lifting platform commonly used in theatre technologies. The configuration of objective device is depicted in Fig. 1. The lifting process is realized in two ways: i) Spiralift, ii) erective chain, so called Serapid. The longitudinal and lateral guiding of the tables is realized by combined bearings connected with tables together with leading U-profiles firmly fixed to the pillars. The aim is determination of modal properties \u2013 mode shapes and natural frequencies. Since the connection of both tables and guiding profiles does not allow the transfer of tension forces in longitudinal direction, the connection is considered as nonlinear",
            " The device was loaded by the mass element located in the middle of both tables \u2013 symmetrical loading of 1000kg was assumed. The model contained 12726 finite elements and 33662 nodes. The two types of materials were assumed in the model \u2013 construction steel E=210000MPa. \u00b5=0.3 and wood E=10000MPa. \u00b5=0.2 which is located in the form of wooden board on the top sides of both tables. The fundamental part of the solution is the realization of nonlinear joint between the tables and guiding U-profiles. The physical realization of connection of tables and guiding profiles is obvious from Fig. 1 and Fig. 3. The tables are provided on both sides by two rollers (Fig. 3) guided in Uprofiles. So the longitudinal and lateral guidance is ensured. The lateral vibration is avoided due to the side areas of U-profile. In case of longitudinal vibration only the compression force can be transferred because the contact between the roller and U-profile occurs only at one its side. This fact is to be considered in FE model. The transfer of the forces between roller and guiding profile was realized by coupling of degrees of freedom belonging to both parts"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001531_s00707-013-0995-y-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001531_s00707-013-0995-y-Figure2-1.png",
        "caption": "Fig. 2 Elastic beam in a moving reference frame",
        "texts": [],
        "surrounding_texts": [
            "(EI: bending stiffness (no shear), c: mass center distance in the undeformed state, g: gravitational constant, gravity acting in xo-direction).\nThe underlined part of Eq. (1) is easily interpreted as the well-known Rayleigh beam equation: for (\u2202m/\u2202x) = \u03c1 A = const, (\u2202 J/\u2202x) = \u03c1 I = const, EI = const (homogeneous uniform beam), one has \u03c1 Av\u0308 \u2212 \u03c1 I v\u0308\u2032\u2032 + EIv\u2032\u2032\u2032\u2032 = 0 [21]. The remainder arises from the \u201cguidance motion,\u201d i.e., the acceleration of the reference frame, and from gravity, respectively.\nOne should have in mind that the partial equations of motion always go along with boundary conditions, so to say the partial differential equations (PDEs) and the boundary conditions (BCs) form an intrinsically tied entity. Having a look at the coupling point x2 = 0, the BCs for v2(0, t) are quite simple (the forearm is clamped at the elbow joint). However, when considering the upper arm (index 1) at x1 = L1, one has\n\u2212 [ \u2202\n\u2202x1\n( [EI]1 v\u2032\u2032 1 )] L + \u2202 J1 \u2202x1 v\u0308\u2032 1L + \u2202 J1 \u2202x1 \u03b3\u03081 + [ m2(L1 + c2 cos \u03b32) \u2212 \u222b B2 v2dm2 sin \u03b32 ] \u03b3\u03081\n\u00d7 [ m2c2 cos \u03b32 \u2212 \u222b B2 v2dm2 sin \u03b32 ] \u03b3\u03082 + m2v\u03081L + ( m2c2v\u0308 \u2032 1L + \u222b B2 v\u03082dm2 ) cos \u03b32\n\u2212 [ m2c2 sin \u03b32 + \u222b B2 v2dm2 cos \u03b32 ] (\u03b3\u03071 + \u03b3\u03072) 2 \u2212 2 [( m2c2v\u0307 \u2032 1L + \u222b B2 v\u03072dm2 ) sin \u03b32 ] (\u03b3\u03071 + \u03b3\u03072) + m2(L1v \u2032 1L \u2212 v1L )\u03b3\u0307 2 1 + m2g(sin \u03b31 + v\u2032 1L cos \u03b31) = 0. (2)\nHere, \u222b\nB2 dm2 denotes integration w.r.t. body 2, i.e., \u222b L2 o (\u2202m2/\u2202x2)dx2.\nThe underlined part of Eq. (2) is easily interpreted as one of the well-known Rayleigh beam BCs (i.e., the transverse force at the beam\u2019s end): (\u2202 J/\u2202x) = \u03c1 I = const, EI = const yields \u03c1 I v\u0308\u2032 \u2212 EIv\u2032\u2032\u2032 = 0. The remainder arises from the \u201cguidance motion\u201d and from forearm reactions.\nThe important role of the BCs is illustrated by the history of plate vibrations: Chladni [8] demonstrated the vibrations experimentally and inspired S. Germain for the mathematical analysis. With some assistance by J.L. de Lagrange, she formulated the PDE in 1811 [12] and the corresponding BCs two years later (and, in revised form, in 1815). However, these were incorrect as Kirchhoff [17] found out, leading to either an infinite number of equilibrium positions or to none. Kirchhoff\u2019s equations, on the other hand, yield an analytical solution only for special BCs, e.g., for rectangular pinned edges (Voigt [27]). It should take some decades until Ritz [23] found solutions for the general case in the form of convergent series expansions. By then, starting with Chladni, it took about 120 years to come to a preliminary end. However, one should also observe its mathematical derivation itself: Kirchhoff\u2019s treatise, with its 37 pages, does obviously not yet represent the easiest and thus purposive procedure. An operator-based method would break down the effort tremendously and need about half a page only [5].",
            "Thus, coming back to Eqs. (1) and (2), the questions arise: how to obtain them? One could think of using Hamilton\u2019s Principle as commonly done. However, such a procedure will result in a nearly unbearable task. We are instead looking for a direct procedure. And, from the engineering point of view: How to solve them? Clearly, Eqs. (1) and (2) only represent part of the problem (we will discuss them in more detail in the \u201cAppendix\u201d). But they already give an impression of the complexity of problems we have to expect.\n2 Single beam, moving reference frame\nThe basis of the following investigations is a single beam in a moving reference frame. The beam element is characterized by dm (mass of a \u201cbeam slide\u201d) and its moment of inertia dJ . In order to simplify matters, we restrict ourselves to plane motions (for the general case, see [7]). The momenta of the beam element are then dpx and dpy where (x,y) represent the motion plane. The angular momentum is dLz , perpendicular to the motion plane. Along with the corresponding mass center velocities y\u0307c, one obtains\n\u239b \u239c\u239d dpx dpy\ndLz\n\u239e \u239f\u23a0 = \u23a1 \u23a2\u23a3 dm 0 0 0 dm 0\n0 0 dJ\n\u23a4 \u23a5\u23a6 y\u0307c := dp, (3)\nall terms defined in a moving frame representation. The mass center velocities are calculated with the velocities of the frame center o, (vox , voy, \u03c9oz), the bending deflection v(x, t), and the bending angle v\u2032(x, t), where the prime indicates spatial derivation. The representation is furthermore augmented with the curvature v\u2032\u2032 (which is later needed for elastic potential forces). All these components are gathered in a vector y\u0307 (the \u201cdescribing velocity\u201d):\ny\u0307c = \u239b \u239d vcx\nvcy \u03c9cz\n\u239e \u23a0 = \u23a1 \u23a3 1 0 \u2212v 0 0 0\n0 1 x 1 0 0 0 0 1 0 1 0\n\u23a4 \u23a6 \u239b \u239c\u239c\u239c\u239c\u239c\u239d\nvox (t) voy(t) \u03c9oz(t) v\u0307(x, t) v\u0307\u2032(x, t) v\u0307\u2032\u2032(x, t)\n\u239e \u239f\u239f\u239f\u239f\u239f\u23a0 := F y\u0307. (4)\nUsing Eq. (4) to formulate the virtual work yields\n\u222b B \u03b4yT c (dp\u0307 + \u03c9\u0303o dp \u2212 df) = 0 = \u222b B \u03b4yT [dMy\u0308 + dGy\u0307 + dKy \u2212 dQ ] . (5)\nHere, \u03c9\u0303o is the skew symmetric spin tensor assigned to the angular velocity \u03c9o of the reference frame.\u222b B indicates \u201cintegration over the body,\u201d i.e., for uniaxial extension in x-direction one has \u222b B dMy\u0308 =\u222b\nL(dM/dx)y\u0308dx , L: beam length. Notice that y\u0308 depends on t and on x. The system matrices read dM =",
            "FTdiag{dm, dm, dJ } F, etc., with F from Eq. (4):\ndM = \u23a1 \u23a3dM11 dM12\ndMT 12 dM22\n\u23a4 \u23a6= \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 dm 0 \u2212dm v 0 0 0 0 dm dm x dm 0 0\n\u2212dm v dm x dJ o dm x dJ 0 0 dm dm x dm 0 0 0 0 dJ 0 dJ 0 0 0 0 0 0 0\n\u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 , (6)\ndG = \u23a1 \u23a3dG11 dG12\ndG21 dG22\n\u23a4 \u23a6= \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 \u2212dm \u2212dm x \u22122dm 0 0 dm 0 \u2212dm v 0 0 0\ndm x dm v 0 2dm v 0 0 dm 0 \u2212dm v 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0\n\u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u03c9oz, (7)\ndK = \u23a1 \u23a3dK11 dK12\ndKT 12 dK22\n\u23a4 \u23a6= \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 \u222b L\nx\n( d f (o) x /d\u03be ) d\u03be 0\n0 0 0 0 0 EIz\n\u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 dx, (8)\ndQ = \u23a1 \u23a3dQ1\ndQ2\n\u23a4 \u23a6= \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 1 0 0 0 1 0\n\u2212v x 1 0 1 0 0 0 1 0 0 0\n\u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u239b \u239d d fx d fy dMz \u239e \u23a0. (9)\nMatrix dK comprises the effect of zero-order forces [w.r.t. the deflections, indicated by superindex (o)] in order to focus the attention on the necessity to formulate the corresponding deviations up to the second order to guarantee correct linearization (sometimes called \u201cdynamical stiffening\u201d).\n2.1 Rigid body motion suppressed\nNeglecting (vox , voy, \u03c9oz), vector y reduces to yT = (v, v\u2032, v\u2032\u2032). The corresponding virtual work then reads\u222b B \u03b4yT[dM22y\u0308 + dK22y] = 0 (for simplicity, dQ2 is assumed zero):\n\u222b B \u03b4yT \u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 \u23a1 \u23a3 \u2202m/\u2202x 0 0 0 \u2202 J/\u2202x 0 0 0 0 \u23a4 \u23a6 \u239b \u239d v\u0308 v\u0308\u2032 v\u0308\u2032\u2032 \u239e \u23a0+ \u23a1 \u23a2\u23a2\u23a2\u23a3 0 0 0 0 L\u222b x ( d f (o) x /d\u03be ) d\u03be 0\n0 0 EIz\n\u23a4 \u23a5\u23a5\u23a5\u23a6 \u239b \u239d v\nv\u2032 v\u2032\u2032 \u239e \u23a0 \u23ab\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23ad dx\n:= \u222b B \u03b4yTdQc = 0. (10)\ndQc is introduced here for abbreviation. As we are interested in q = v for minimal representation, we may define a (spatial) operator D ,\nyT = (v v\u2032 v\u2032\u2032)T = ( 1 + \u2202\n\u2202x + \u22022 \u2202x2\n)T\n\u25e6 v := D T \u25e6 v, (11)\nwhich immediately leads to the partial differential equations via a new operator D . D is the same as D except that the odd derivatives change their sign. Simultaneously, one obtains the operators Bi for the boundary"
        ]
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.23-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.23-1.png",
        "caption": "Fig. 3.23",
        "texts": [
            " Then we obtain from (3.30b) with x\u0308A = y\u0308A = 0 and \u222b (\u03be2 + \u03b72)dm = \u0398A \u0398A \u03d5\u0308 = MA . (3.33) This is the result that we already found in the case of a rotation about a fixed axis (see Section 3.2.1). The axis is perpendicular to the x, y-plane and it passes through point A. In this case, the point of reference in the principle of angular momentum may either be the center of mass C or the fixed point A. As an illustrative example we consider a homogeneous solid sphere that moves down a rough inclined plane (Fig. 3.23a). We first assume that the sphere rolls without slipping. The free-body diagram (Fig. 3.23b) shows the forces acting on the sphere. The coordinates that describe the motion are the position xc of the center of the sphere and the angle \u03d5 of rotation. With y\u0308c = 0, the principles of linear and angular momentum yield \u2198 : mx\u0308c = mg sin \u03b1\u2212H, (a) \u2197 : 0 = N \u2212mg cos\u03b1 \u2192 N = mg cos\u03b1, (b) C: \u0398C \u03d5\u0308 = r H. (c) The moment of inertia of the sphere is given by \u0398C = 2 5 mr2 (see Example 3.5). Since it is assumed that the sphere rolls without slipping, the kinematic relation x\u0307c = r\u03d5\u0307 \u2192 \u03d5\u0308 = x\u0308c r (d) holds (see (3",
            " Thus, (a), (c), and (d) yield the acceleration of the center of mass C: mx\u0308c = mg sin\u03b1\u2212 \u0398C r2 x\u0308c \u2192 x\u0308c = g sin \u03b1 1 + \u0398C mr2 = 5 7 g sin \u03b1 . The friction force during rolling follows from (a): H = m(g sin\u03b1\u2212 x\u0308c) = 2 7 mg sin \u03b1 . Now we are able to formulate the condition which must be satisfied by the coefficient of static friction \u03bc0 in order to ensure rolling of the sphere: H \u03bc0N \u2192 \u03bc0 H N = 2 7mg sin \u03b1 mg cos\u03b1 = 2 7 tan \u03b1 . If \u03bc0 does not satisfy this condition, the sphere will slip on the inclined plane. Then the friction force H has to be replaced by the dynamic friction force R in Fig. 3.23b and in (a) and (c): mx\u0308c = mg sin \u03b1\u2212R, N = mg cos\u03b1, \u0398C \u03d5\u0308 = r R . (e) In addition we have to use the friction law R = \u03bcN . (f) When the sphere is slipping, no relation exists between x\u0308c and \u03d5\u0308: they are independent of each other. From (e) and (f) we obtain the accelerations x\u0308c = g(sin \u03b1\u2212 \u03bc cos\u03b1), \u03d5\u0308 = 5 \u03bcg 2 r cos\u03b1 . E3.9 Example 3.9 A simplified model of a car is shown in Fig. 3.24a. It consists of a rigid body (weight W = mg, center of mass at C) and massless wheels. Find the maximum acceleration of the car on a rough horizontal surface (coefficient of static friction \u03bc0), if the engine only drives a) the rear wheels (the front wheels are rolling freely), b) the front wheels (the rear wheels are rolling freely)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002989_2425296.2425316-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002989_2425296.2425316-Figure6-1.png",
        "caption": "Figure 6: Sphere-sphere collision detection.",
        "texts": [
            " The representation of the robot by the spherical bounding volumes is shown in Figure 5. Although the robot can be modeled more precisely by increasing the number of elements, it takes a longer computational time to calculate the distances between the numerous elements for collision detection. In this paper, collision detection is only done for two cases, which are (i) between the upper arm and torso, and (ii) between the forearm and the torso. It is assumed that the joint limit of the elbow prevents the upper arm and the forearm from coming into contact. In Figure 6, the critical distance dij is the distance between the centres of the two spheres. When dij > ri + rj , then there is no risk of collision; otherwise, the two spheres are in collision. When two spheres are in contact each other, a virtual repulsive force for collision avoidance is generated between them. Referring to Figure 6, the direction of the virtual force is directed along the line passing through the centres of the two spheres and the magnitude of the force F is shown below: Fij = K|dij \u2212 (ri + rj)| (6) where K is a spring constant. During collision of robot segments, consider the case of a repulsive force Fr1 being generated on the upper arm at a distance of lr1 from the shoulder joint as shown in Figure 7. The force Fr1 can be transformed into an equivalent force at the elbow using the expression below: Felb = Lr1 L1 Fr1 (7) Similarly, a repulsive force Fr2 generated on the forearm at a distance of lr2 from the elbow joint can be transformed into an equivalent force at the wrist by: Fwr = Lr2 L2 Fr2 (8) For multiples collisions between bounding spheres, Felb and Fwr become a summation of the equivalent forces at the elbow and wrist respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002529_mechatron.2011.5961097-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002529_mechatron.2011.5961097-Figure3-1.png",
        "caption": "Fig. 3. Multi-phalanxes gripper with different lengths of links As a result of appearance of these forces, the load on the surface, created by supporting phalanxes, is increased.",
        "texts": [
            " Then, for various numbers of phalanxes of GD we will obtain the following picture, see Fig.2. Analyzing the three-phalanxes gripper, one can conclude, that in general it will be ineffective, since at definite dimensions of the sphere (this range R\u2265a*sin\u03b1 is rather wide) one phalanx does not participate in fixation of the object, not contacting with its surface. However, even such gripping will be enough in order to fix safely and position accurately the manipulated object, if the pressure, provided by phalanxes, does not exceed the maximum allowable one. The scheme, shown in Fig. 3, is characterized by the presence of one long phalanx. It provides the simplest control and the least stresses, appearing on the surface of the object during gripping. It is achieved due to the projection of vectors of the load from the phalanx of GD on the surface of the object. Multi-link (with the same length) gripping devices create the stresses, directed oppositely from useful forces, by means of their first phalanxes. \u2013 projection of the reaction on the axis \u0443, \u2211 \u2013 projection of the inverse reaction on the axis \u0443, \u2211 \u2013 projection of the useful force on the axis \u0443. The scheme of the layout, represented in Fig. 3, solves this problem. Since the arm \u0430\u0430\u2019 is big, the stress in the contact point will be small. Forces, appearing at the point P2, will be less than forces, appearing at the point P1, if the condition of torques equality is satisfied. Let us consider the loads in the moment of gripping the object: \u041c \u2013 mass of the weight; \u041c\u043a \u2013 limiting torque in the joint, when the drive of the phalanx stops increasing the load; R \u2013 radius of the sphere; \u0430\u0441 \u2013 distance between centers of the sphere and gripper. Here, the limiting factors will be: P\u043c\u0430\u0445 \u2013 maximum pressure on the surface of the object; M\u043a_\u043c\u0430\u0445 \u2013 maximum torque in the joint of the finger"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002526_kem.572.397-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002526_kem.572.397-Figure2-1.png",
        "caption": "Fig. 2 The basic parameters of the specimen",
        "texts": [
            "150, Rice University, Fondren Library, Houston, USA-19/05/15,08:46:37) Friction Specimens. Sliding friction couples are made up of the upper specimen and downbeat specimen, the upper one keeps fixed and the downbeat one rotates. Specimens are designed and made in accordance to the National Standard. The width of the underneath specimen which simulates the roller of the sliding bearing made of 40Cr with the hardness HRC58-62, is 10mm. The width of the upper one made of ZCuPb10Sn10 with hardness HBS70, is 4mm. Structure of the specimen is shown as Fig.2. Lubrication Method. The oil-air lubrication is adopted. The air pressure is 0-0.5Mpa. The L-TSA46 machine lubricating oil supply volume controlled exactly by adjusting the time of oil supply is 0-50ml/h. Lubricating oil and air are transported from oil-air pipe through the nozzle to the lube point. Oil-air nozzles are direct injection nozzle without swirl chamber and conical nozzle with swirl chamber shown as Fig.3 and Fig.4. The rotation speed of the test is 210rpm with temperature of laboratory 17"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003474_amm.130-134.1205-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003474_amm.130-134.1205-Figure2-1.png",
        "caption": "Fig. 2 Structure of the measuring head",
        "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. (ID: 129.128.216.34, University of Alberta, Edmonton, Canada-27/04/15,03:58:33) dkd rL VC V m SS \u22c5=\u22c5= \u03c0\u03b52 0 (1) The equation (1) can be obtained approximately when the capacitive sensor locates in centre of the measured hole. k is the equivalent coefficient for the instrument. Output voltage is linear with diameter of measured hole. The mm.01\u03c6 capacitive sensor is designed to satisfy smaller inner diameter measurement. And structure of the measuring head is shown in Fig.2: The measuring system composes of a main measuring bench, stepping motor control module, vision module and a two-channel capacitive micrometer (circuit processing module). The mechanical structure of the system is shown in Fig.3: In this automatic measuring system, accurate positioning is achieved by cooperating stepping motors with gratings respectively in X, Y, Z direction. Stepping motor is driven through PCI port by computer and the indication of grating is regarded as effective information to control the stepping motor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.55-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.55-1.png",
        "caption": "Fig. 3.55",
        "texts": [
            "31 A homogeneous triangular plate of weight W = mg is suspended from three strings with negligible mass. Determine the acceleration of the plate and the forces in the strings just after string 3 is cut. Results: see (A) a = g sin \u03b1, S1 = mg 6 (sin \u03b1+4 cos\u03b1) , S2 = mg 6 (\u2212 sin \u03b1+2 cos\u03b1) . E3.32 Example 3.32 A sphere (mass m1, radius r) and a cylindrical wheel (mass m2, radius r) are connected by two bars (mass of each bar m3/2, length l). They roll down a rough inclined plane (with angle \u03b1) without slipping (Fig. 3.55). Find the acceleration of the bars. Result: see (B) a = (m1 + m2 + m3)sin \u03b1 7 m1/5 + 3 m2/2 + m3 g . 3.5 Supplementary Examples 203 E3.33Example 3.33 The cylindrical shaft shown in Fig. 3.56 has a varying mass density given by \u03c1 = \u03c10(1 + \u03b1r). Find the moments of inertia \u0398x and \u0398y. x R r y y z l Fig. 3.56 Results: see (A) \u0398x = \u03c0 2 \u03c10lR 4 ( 1 + 4 5 \u03b1R ) , \u0398y = \u03c0\u03c10R 2l [R2 4 + l2 3 + \u03b1 (R3 5 + 2 9 Rl2 )] . E3.34Example 3.34 Determine the moment of inertia \u0398a of a homogeneous torus with a circular cross section and mass m"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002280_synasc.2013.9-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002280_synasc.2013.9-Figure2-1.png",
        "caption": "Figure 2. Isosceles triangle \u25b3 GFE constructed by a fold",
        "texts": [
            " The folds by which we make the polygonal knot from the tape are called knot folds. One of the most fundamental properties of the crossing of the knot fold is expressed in the following lemma. Lemma 1 (Isosceles crossing). For any origami ABCD, point E \u2208 AB and point F \u2208 CD, for any point G such that G \u2208 CD \u2227 G \u2208 EAFE, \u25b3 GFE is isosceles with \u2223FG\u2223 = \u2223EG\u2223. The lemma is written in somewhat contrived statement for reasons explained shortly. The meaning is clear if it is accompanied with the illustrative figure (cf. Fig.2). The proof is straightforward by elementary geometrical reasoning, and is omitted here. Note, however, that the lemma holds not only for the configuration of the concerned points as depicted in Fig. 2, but also for the cases of EF\u22a5AB and the cases of E or F being outside the given tape. All the constructions and the propositions given in this paper are treated by EOS. We do not need to handle the \u201cexceptional\u201d cases separately, as the EOS prover checks them automatically. Furthermore we combine the geometrical construction and the verification of the properties that we claim by the construction. In all the examples of this paper, the EOS prover uses Gro\u0308bner bases method [7]. We take arbitral points E and F and an arbitrary rectangular origami, and further assume, without loss of generality, that the height of the origami is of unit length, and compute Gro\u0308bner bases of the set of the polynomials generated from the logical specification of the propositions, over the coefficient domain of rational functions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002837_2011-01-0197-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002837_2011-01-0197-Figure9-1.png",
        "caption": "Figure 9. Neutral fiber location in the bearing",
        "texts": [
            " As the load direction is coincident with the conrod axes, there is no significant housing elongation and by the way, no interference fit modification. For a bearing with a radial clearance Cr, and an inner radius Ri, we can estimate the stress in the overlay bearing. Thanks to the slippage between the bearing and the housing, the stress generated by the housing wrapping is the normal stress \u03c3x. With bending relationships for curved beam from \u2018strength of material\u2019 theory, the stress tensor is: (eq.2) with ynf the location of the considered fiber in the layer with respect to the neutral fiber of the global bearing (figure 9). The neutral fiber location with respect to the symmetrical thickness is : For a quick evaluation, we can consider that the location of the neutral fiber is ynf =tbl/2 with tbl the bearing thickness. Local housing distortion stress In the crown conrod area where our analysis is focused, the local distortion, estimated with E.H.L. simulation [10], is negligible. In the axial bearing direction, the housing bore distortion is smaller than 0.3 \u00b5m for the highest loaded bearing (Pressure max: 250 MPa)",
            " This extra diameter is called \u0394D. During bearing assembly process, the diameter reduction is due to force reaction at the edges of the housing. These reactions generate bending stresses in the bearing. The stress are permanent in the substrate. For the overlay, the free spread stress disappears with the relaxation effect. With \u2018Strength of Material\u2019 theory, we simply obtained the following relationship for the stress tensor in the substrate: (eq.4) In this relationship, \u03b8 is measured from the joint face of the bearing (figure 9) and ynf is evaluated from the neutral fiber [13]. Fitting friction stress During the fitting bearing process, friction shear stress is generated at the interface between the bearing and the housing. Nevertheless, as the housing is distorted during each engine cycle, the friction stress due to the fitting process vanishes quite rapidly. As we are looking for endurance behavior after high number of cycles, it is not necessary to consider this kind of stress. Global thermal stress This is the most important contributor in terms of stress"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.100-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.100-1.png",
        "caption": "Fig. 2.100 4PaRRRR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology Pa\\R\\R||R\\R (a) and Pa||R||R\\R||R (b)",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.21-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.21-1.png",
        "caption": "Figure 4.21 Spray lubrication: (a) spraying in the starting side of teeth mesh, (b) spraying in the end side of teeth mesh",
        "texts": [
            " Between the cooler and the drive a bypass valve (8) is fitted through which the whole of the oil, or a part of it can get back into the tank, bypassing the driver. Thus the oil temperature in the system is regulated. Of course, thermometers (9) are mounted before and after the coolers, as well as the tank vent valve (10) on the top, the level gauge (11) outside the tank and a valve (12) on the bottom for draining oil. Under the pressure of the pump, the oil is sprayed in the gear mesh by means of a nozzle (sprayer). When the peripheral speed v\u00bc 10\u201360m/s, oil is sprayed at the starting side of the teeth mesh (Figure 4.21) while at very high peripheral speeds, oil is sprayed at the end side of the teeth mesh (Figure 4.21b). In the latter mode, the crash speed of sprayed oil with the teeth is so large (equal to the sum of their individual speeds) that the oil cannot remain on the teeth flanks, but is converted into an oil mist that fills the whole inside of the drive and always drops a little on all of the teeth and other places. In the spray lubrication, the quantity of a lubricant in the housing should be at least equal to the flow of oil in 20min of work. The replacement of a lubricant is determined in each case with reference to the operating conditions, the environment and other factors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure19-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure19-1.png",
        "caption": "Figure 19. Normalised Stress at RSRB region (Contact analysis)",
        "texts": [],
        "surrounding_texts": [
            "Surface to surface contact is defined for the following contact pairs \u2022 RSFB bracket & External flitch \u2022 External Flitch & FSM \u2022 FSM & Internal flitch \u2022 Internal Flitch & RSFB Cross member"
        ]
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure14-1.png",
        "caption": "Figure 14. Input Force to the Scapulothoracic Joint",
        "texts": [
            " SHOULDER FORCE WHEN EXTERNAL INPUT IS GIVEN The force of shoulder muscle is analyzed when three different outside forces are given. A. Input to Articulatio Acromioclavicularis (Analysis 1) External input is added at the articulation sternoclavicularis (Fig. 12 and Table V). a F represents arm\u2019s weight. b F represents the external input. The force of shoulder muscle, and the force of clavicle and the rib are shown in Fig. 13. B. Input to Scapulothoracic Joint (Analysis 2) External input c F is added at the scapulothoracic joint (Fig. 14 and Table VI). The force of shoulder muscle, the force of clavicle and the rib are shown in Fig. 15. C. Input to Articulatio Acromioclavicularis and Scapulothoracic Joint (Analysis 3) External input is added evenly at the articulation sternoclavicularis ( b F ) and the scapulothoracic joint ( c F ) (Fig. 16 and Table VII). The force of clavicle and the rib are shown in Fig. 17. D. Discussion In order to minimize the burden on the shoulder muscles, we compared the amount of force exercised by the muscles in the analysis 1 to 3 by integrating the absolute values of muscles forces 1 F , 2 F , 3 F , 4 F , when 2 \u03b8 is varied from 80 degrees to 110 degrees, as this is supposed to be the range most used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003485_s1068798x12060093-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003485_s1068798x12060093-Figure1-1.png",
        "caption": "Fig. 1. Model of turbine rotor: (1\u20134) positions of radial safety bearings; (5) position of axial safety bearings; (6) generator rotor; (7) elastic coupling; (8) turbocom pressor rotor.",
        "texts": [
            " To this end, we analyze the influence of the intrinsic forms of rotor vibration on its dynamics as it passes through resonant zones in coasting. The rotor system is characterized by a vertical position of the turbine rotor; considerable mass (around 23 t); total height of around 25 m and rotor diameter at the safety bearing positions of around 0.5 m [5]; and res onant zones up to the rotor\u2019s working speed. The tur bine rotor in fact consists of two rotors: that of the gen erator and that of the turbocompressor, which are elas tically coupled (Fig. 1). Each rotor is suspended on electromagnetic bearings: two radial bearings and one axial bearing. DOI: 10.3103/S1068798X12060093 RUSSIAN ENGINEERING RESEARCH Vol. 32 No. 7\u20138 2012 SIMULATION OF FLEXIBLE TURBINE ROTORS 533 Satisfactory performance of the safety bearing under the action of the rotor depends on minimization of the load on the supporting structures. The motion and interaction of the vertical rotor with the radial safety bearings may take different forms (impact, slip, or rolling over the internal bearing race, precession)",
            " The model incorporates the basic factors that determine rotor motion in the safety bearing: \u23afthe local geometry\u2014in particular, the gap between the rotor and the safety bearing and the path of the safety bearing\u2019s shoe; \u23afthe initial conditions in terms of the rotor\u2019s speed and its position with respect to the components of the safety bearing; \u23afthe rigidity of the rotor sections (including flange joints); \u23afthe gyroscopic torques of the rotor and sus pended masses; \u23afthe distribution of the residual imbalance over the length of the rotor; \u23afdamping in the rotor\u2013bearing system: specifi cally, the internal friction at contacts of the rotor com ponents; the external friction of the rotor and compo nents of the safety bearing; and the friction within components of the safety bearing; \u23afthe rigidity of the contact surfaces of the rotor and shoe and the rigidity of contact surfaces between components of the safety bearing; \u23afthe elasticity of the key components of the safety bearing; \u23afthe braking torque of the rotor system in coasting. In the simulation, each turbine rotor (Fig. 1) is rep resented by a chain of solid cylinders coupled by elas tic elements. The mass, inertial, and geometric char acteristics of the model match those of the turbine rotor; gyroscopic torques may also be taken into account. To refine the model parameters, optimiza tion ensures that the forms and eigenfrequencies of flexural vibrations correspond to the 3D finite ele ment model of the turbine rotor. The frequency differ ence of the vibrations of the model consisting of cylin ders and elastic elements and the 3D finite element model is 4%"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure45-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure45-1.png",
        "caption": "Fig. 45 CoSIMT Toe/Heel chamfering with different blades\u2014straight bevel gears",
        "texts": [
            " The main reason is tool size which must be limited in order not to damage the tooth flank behind the tooth Toe or Heel edge; this therefore limits the CoSIMT diameter and the chamfer angle. On the other hand, the sturdiness of CoSIMT tools ensure a long service life and makes them choice tools. In addition, while chamfering the left tooth edge with the CoSiMT\u2019s outside blade\u2014i.e. the blade on the opposite side of the arbor\u2014causes no risk of interference between the tool\u2019s arbor and the work piece (left, Fig. 45), using the inside blade\u2014 the blade on the same side as the arbor\u2014requires an elongated arbor (right, Fig. 45) to avoid interference and this is likely to lead to vibrations and uneven results. This implies that, for practical reasons, the CoSIMT outside blade should be used to chamfer both the left and right tooth Toe and Heel edges, as shown in Fig. 46. This also implies that the process is not continuous: all the Toe/Heel edges on one tooth flank are done, and then all the edges of the other tooth flank. In practice, this does not involve excessive cycle time and is well worth given all the chamfering can be done using a single tool"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003296_010-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003296_010-Figure1-1.png",
        "caption": "Figure 1. The object is launched from the \u2018North pole\u2019. The Earth has radius RE, and the apogee [\u2018peak\u2019] has radius rmax. The angle \u03c60 characterizes both the launch and the fall back to Earth, due to reflection symmetry with respect to the major axis of the ellipse. The total angular displacement is 2 \u03b8 .",
        "texts": [
            " The escape speed is given by \u03bd\u221e = \u221a 2 \u03bd. We consider the case when the launch speed is arbitrarily large, 0 < v0 < \u03bd\u221e. We ignore the effects due to the rotation of the Earth and air resistance. We assume without loss of generality that the object is launched from the \u2018North pole\u2019. If the total mechanical energy is less than zero, E = T + V < 0, the object is bound to the Earth, and the orbit is an ellipse with the centre of the Earth as one of its focal points. The major axis of the ellipse is rotated by the angle \u03b8 , cf figure 1. The \u2018peak\u2019 of the trajectory is defined by vanishing radial speed vr = r\u0307 = 0. The peak of the trajectory lies on the major axis: the orbit is symmetrical and the object falls back to the Earth when its total angular displacement is 2 \u03b8 . It follows that the range is given by R = 2RE \u03b8 . The case of horizontal launches requires special consideration for 0 < v0 < \u03bd and \u03bd < v0 < \u03bd\u221e. The outline of the paper is as follows. In section 2, we summarize results from orbital dynamics for an object in a central potential",
            " We have the conservation of energy e = 1 2 ( r\u03072 + r2\u03b8\u03072 ) + (r), where e = E/m is the energy per mass, and = V/m is the gravitational potential. The conservation of angular momentum is written as r2\u03b8\u0307 = l, where l = L/m is the angular momentum per mass. This gives the angular speed \u03b8\u0307 = l/r2, so that the radial velocity follows r\u03072 = 2[e \u2212 (r)] \u2212 l2/r2. The constants of motion are determined by the launch parameters, i.e. the initial radius RE, the speed v0, and the angle \u03c60. The launch parameters are shown in figure 1. We find for the angular momentum l = v0 cos \u03c60RE, (2) and for the total energy e = 1 2v2 0 + (RE). (3) We thus have r\u03072 = v2 0 + 2[ (RE) \u2212 (r)] \u2212 l2/r2. The maximum radius rmax follows from r\u0307 = 0 or v2 0 + 2[ (RE) \u2212 (rmax)] \u2212 l2/r2 max = 0. We assume spherical symmetry, such that the gravitational potential is given by (r) = \u2212GME/r , where ME is the mass of the Earth and G is the universal gravitational constant. The inverse-square law does not have an intrinsic length scale so that the Earth\u2019s radius RE defines the dimensionless radius r/RE , and \u03bd = \u221a gRE is the characteristic speed for the problem"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003848_imece2012-87624-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003848_imece2012-87624-Figure6-1.png",
        "caption": "FIGURE 6. TOP AND FRONT VIEW OF KAI MANIPULATOR IN FIRST EIGENMODE AT DIFFERENT POSES",
        "texts": [
            " The current pose of the manipulator is described by the joint angles \u03b1\u0302. With yII = y\u0307I and neglecting internal damping, (15) can be rewritten as a second order linear parameter dependent system M\u0302(\u03b10, \u03b1\u0302)y\u0308I \u2212 K\u0302(\u03b10, \u03b1\u0302)\u2206yI = TT0. (17) The turntable of the manipulator is driven by a hydraulic system of high internal stiffness, therefore the dynamics of the turntable \u03b10 can be neglected. Within the topic of this paper, the turntable is assumed to be clamped \u03b1\u03070 = 0, which can be included into (17) by model reduction techniques. As shown in Figure 6, the first eigenmode of the KAI manipulator represents a similar bending of the flexible elements in the horizontal plane. However the composition of shape functions within each eigenmode, as well as the corresponding eigenfrequency depends on the pose of the manipulator. The tree different exemplary poses are shown in Figure 3. In Table 4 the eigenfrequency of the first mode is listed at each exemplary pose. The shape function coefficients of the first two shape functions of each flexible link element are shown, qi1 corresponds to the the shape function \u03a6i1, which is the torsion mode of the body Ki and qi2 denotes the coefficient of the first horizontal bending mode of Ki"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.18-1.png",
        "caption": "Figure 12.18 A suspended disc on a needle.",
        "texts": [
            " Therefore, the angular velocity vector B G\u03c9B uniformly sweeps a constant cone with the z-axis as the axis of symmetry. Depending on I3 and I1, the body can turn faster or slower than B G\u03c9B . For a flat disc with I3 >I1, we find that the \u03b8 -axis rotates faster than the x-axis in the same direction, and for an elongated body with I3 < I1, the \u03b8 -axis rotates slower than the x-axis in the opposite direction. For I3 = 2I1, we have = \u03c93. The special case I3 = I1 provides = 0, and therefore, the axis of angular velocity B G\u03c9B remains motionless. Example 740 A Rotating Disc on a Needle Figure 12.18 illustrates a homogeneous disc of mass m and radius R that is suspended at its center on a needle. We attach a body frame B at the center of the disc such that the z-axis is perpendicular to the disc and the (x, y)-plane is the face plane of the disc. We give an initial rotation G\u03c9B = \u03c90 about an axis that makes an angle \u03b1 with the z-axis, looking for the consequent motion of the disc. The mass moment matrix of the disc is BI = \u23a1 \u23a2 \u23a2 \u23a3 1 4mR2 0 0 0 1 4mR2 0 0 0 1 2mR2 \u23a4 \u23a5 \u23a5 \u23a6 (12.345) Let us assume that the initial angular velocity in the (y, z)-plane is B G\u03c9B (0) = \u03c90 sin \u03b1 j\u0302 + \u03c90 cos \u03b1 k\u0302 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002825_isie.2013.6563881-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002825_isie.2013.6563881-Figure2-1.png",
        "caption": "Fig. 2 Model of a two-pole asynchronous machine with a 28 bars squirrel cage rotor.",
        "texts": [
            " Thus, the dimension of the matrix is determined by the number of stator windings and the rotor bars.               r Nr 12 12 Nr Nr r Nr r Nr r Nr 3 N cos( ) cos( ) [L ] L cos 4 / 3 cos 4 / 3 cos 2 / 3 cos 2 / 3 cos (N 1) cos 4 / 3 (N 1) cos 2 / 3 (N 1)                                                 (12) 12L and 2L are calculated based on the construction and geometry of the machine. The angle \u03b3 is measured from the magnetic axis of the stator coil U to the first mesh of the rotor as depicted in Fig 2. The model presented above is well-known in the standard literature [13]. Its main drawback is the higher order of the system of differential equations and the fact that the parameters of the cage have to be calculated based on the geometry of the rotor and are not easily available. The great advantage is that it allows including an asymmetry in the calculation of the dynamic behavior of the IM. Therefore, in previous works this model was implemented in a MATLABSimulink environment, the parameters were calculated for the different geometries of rotor cages and finally it was corroborated by laboratory measurements [3]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003035_icra.2013.6630606-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003035_icra.2013.6630606-Figure3-1.png",
        "caption": "Fig. 3. A set of structuring elements for isotropic and anisotropic mobility erosion. The above image shows vehicle reachability under braking maneuvers (assuming a fixed worst-case decceleration). The colors correspond to the minimum initial velocity required to reach any given state under worst case braking maneuvers (white -5 m/s and black - 0 m/s). Each level set in the above images is a unique SE and corresponds to a mobility (velocity or kinetic energy) state. The isotropic SE set ignores vehicle orientation to create conservative estimates of vehicle reachability. The anisotropic SE set is shown for a steered vehicle and it only considers a fixed steering range (\u00b130\u2218) to generate less conservative reachability estimates.",
        "texts": [
            " For a point mass with initial speed limit v (or kinetic energy limit), assuming double integrator dynamics the worst-case stopping distance (d(v)) is given as v2 2amax , where amax is the maximum possible deceleration that can be achieved in the given terrain. Reachability is also affected by system latency ( ) which adds an increment to the stopping distance (v\u00d7 ). In addition to reachability, position uncertainty can be taken into account in the structuring element by dilating it with the 2-sigma uncertainty ellipse. Vehicle reachability with and without steering considerations is shown in Figure 3. Each grid value specifies the initial velocity required to be able to reach that point under a braking maneuver. This reachability map was generated by collating forward simulations of braking maneuvers for a discrete set of starting conditions (set of initial velocities and fixed steering angles). The level sets of the reachability map provide the binary neighborhood required to create a SE for a given initial condition. Alternatively an analytical function can be used when appropriate. This process can be performed offline and the SE set for various initial conditions can be stored in memory. Two sample SE sets are illustrated in Figure 3. The isotropic SE set ignores vehicle orientation to create conservative estimates of vehicle reachability. The anisotropic SE set shown in Figure 3(b) is for a steered vehicle and it only considers a fixed steering range (\u00b130\u2218) for a given position and orientation of the vehicle to generate less conservative reachability estimates. The mobility erosion operation is given in Equation (2) where the SE is adapted according to the current-state mobility values derived from the mobility function (I()). Mobility erosion searches for the optimal speed limit below the prescribed maximum vehicle speed limit(m \u2208 [0, I(s)]) that maximizes the minimum mobility in the neighborhood This leads to a maxmin formulation as shown below in Equation (2)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001783_9781118316887.ch2-Figure2.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001783_9781118316887.ch2-Figure2.4-1.png",
        "caption": "Figure 2.4-2: Graphical representation of electromechanical energy conversion for Xi path B to A.",
        "texts": [
            " [AW) = f J; AWC = f J] SP2.3-4 i = b(x)\\2. Express Wf(X, x) and WC(A, x). [Wf(X, x) = \\b(x)X3; WC(X, x) = lb(x)X3} OF ENERGY CONVERSION Before proceeding to the derivation of expressions for the electromagnetic force, it is instructive to consider briefly a graphical interpretation of the energy conversion process. For this purpose, let us again refer to the elementary system shown in Fig. 2.2-3 and assume that as the movable member moves from x = xa to x = xb, where xb < xa, the M characteristics are given by Fig. 2.4-1. Let us furthermore assume that, as the member moves from xa to xb, the M trajectory moves from point A to point B. The exact trajectory from A to B is determined by the combined dynamics of the electrical and mechanical systems, and any variation in v and / which may occur. Now, the area OACO represents the original energy stored in the coupling field; area OBDO represents the final energy stored in the field. Therefore, the change in field energy is AW) = area OBDO - area OACO (2.4-1) The change in We, denoted as AWe, is AWe = [ * id\\ = area CABDC (2.4-2) From (2.2-6), AWm = AW) - AWe (2.4-3) Hence, by adding and subtracting the appropriate areas in Fig. 2.4-1, we can obtain AWm. In particular, AWm = area OBDO - area OACO - area CABDC = -area OABO (2.4-4) The energy contributed to the coupling field from the mechanical system AWm is negative. Energy has been supplied to the mechanical system from the coupling field, part of which came from the energy stored in the field, and part of which came from the electric system. If the member is now moved back to xa, the \\i trajectory may be as shown in Fig. 2.4-2. Here, AWm is still area OABO, but it is positive, which means that energy was supplied from the mechanical system to the coupling field, part of which is stored in the field and part of which is transferred to the electric system. The energy supplied by the mechanical system during the motion from B to A (area OABO in Fig. 2.4-2) is larger than the energy supplied to the mechanical system during the original motion from A to B (area OABO in Fig. 2.4-1). Therefore, the net energy supplied by the mechanical system for the complete cycle is positive. The net AWm for the cycle from A to B back to A is the shaded area shown in Fig. 2.4-3. Since the coupling field energy at point A is uniquely determined from the mechanical displacement and current at point A, the net change in field energy is zero as we move from A to B and back to A. Since AWf is zero for this cycle, AWm = -AWe (2.4-5) For the cycle shown, the net AWe is negative since the magnitude of the change in We is larger when we went from B to A than from A to B, and AWe is negative from B to A. If the trajectory had been in the counterclockwise direction, the net AWe would have been positive and the net AWm negative",
            "6-2, e/ jumps to 5 V when the source voltage is stepped from zero to 5 V and jumps from zero to \u20145 V when the source voltage is stepped from 5 V to zero. Why? [At first t = 0+, v \u2014 L{x)(di/dt)\\ at second \u00a3 = 0+, L(x)(di/dt) = -ir] SP2.6-2 Consider Fig. 2.6-1 with initial operation at point 2. Determine the final operating value of x (a) if / is stepped from zero to \u2014 1 N; (6) if / is zero but v is stepped from 5 V to 10 V [(a) and (b) x = 0] SP2.6-3 Assume that the elementary electromagnet shown in Fig. 2.2-3 portrays the Xi characteristics shown in Fig 2.4-1. As the system moves from xa to Xb, the Xi trajectory moves from A to 5 , as shown in Fig. 2.4-1. Assume steady-state operation exists at A and B. (a) Does the voltage v increase or decrease? (6) Does the applied force / increase or decrease? [(a) and (b) decrease] SP2.6-4 Why does not i(t) portray an exponential increase when the source voltage is increased from zero to 5 V in Fig. 2.6-2? [L(x)] An elementary two-pole, single-phase reluctance machine, which was first shown in Fig. 1.7-2, is shown in a slightly different form in Fig. 2.7-1. In particular, the notation has been changed, that is, winding 1 is now winding as"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002903_amr.381.81-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002903_amr.381.81-Figure1-1.png",
        "caption": "Figure 1 shows the orthogonal coordinate system in face gear machining process, the angle of face gear and the cutting tool is 90 \u00b0. The cutting tool involutes tooth surface equation in the coordinate system is [5]:",
        "texts": [],
        "surrounding_texts": [
            "Face gear drive is a new transmission type, meshing with a cylindrical gear. It has many advantages, for example the system is light and little on volume, lower noise, and is dominant on kinetic property and preference on separating moment. And face gear has been applied in drive system of helicopters [1]. Litvin [2] finished the static contacting analysis (TCA) on face gear with a 3-D model. Yang Lianshun [3] analyzed the bending stresses behavior of a tooth, and got the position where maximum bending stresses occur and the case of its movement. Guo Hui [4] researched the movement of the contact point and the influence of pressure angle of the bending strength. These references involved some research on bending stresses in a certain extent, lacking of the analysis about bending stresses of face gears with different parameters yet. In this article, UG is used to simulate manufacturing process of face gear, and the face gear model is created. After reverse modeling, the completed model that can be used in ANSYS is finished. And then the rules of bending stresses of face gears with different module, teeth width and number of teeth are researched. Mathematical model of the orthogonal Face Gear             +++\u2212 +\u2212+\u00b1 = 1 ]sin()[cos( )]cos()[sin( ),( 00 00 s ssksssbs sssssbs sss u r r ur \u03b8\u03b8\u03b8\u03b8\u03b8 \u03b8\u03b8\u03b8\u03b8\u03b8 \u03b8 (1) Where rbs is the cutting tool involute radius of base circle; \u00b5s is parameters of cutter tooth surface point on the axial direction; \u03b8s is angle parameters of point in cutter involute line; \u03b80s is angle from cutter alveolar symmetrical line to the involute starting point), \"\u00b1\"symbol correspond to the both sides of cutter alveolar involute . With the principle of the gear meshing, gear meshing equations and coordinate transformation from Ss to S2: 2 2 2 2 2 2 2 2 2 2 cos cos cos sin sin sin 0 cos sin sin cos sin cos sin cos sin sin sin cos 0 cos cos sin cos cos cos sin sin sin cos cos 0 0 0 0 1 s s m m s m s s s m m s m s m s m s m \u03c6 \u03c6 \u03c6 \u03c6 \u03b3 \u03c6 \u03b3 \u03c6 \u03c6 \u03b3 \u03c6 \u03c6 \u03c6 \u03c6 \u03c6 \u03c6 \u03b3 \u03c6 \u03b3 \u03c6 \u03c6 \u03b3 \u03c6 \u03c6 \u03b3 \u03c6 \u03b3 \u03c6 \u03b3 \u2212  \u2212 + +   \u2212 = \u2212  + +         2s M (2) Obtain the following equations:     \u22c5= =\u22c5 ),()(),,( 0 )( 2 )2( 2 )()( 2 ss s ssssss s s s s urMur nv \u03b8\u03c6\u03c6\u03b8 (3) Where \u03bds2 (s) is the relative velocity of tooth surface meshing point, ns (s) is normal to the cutter surface, Ms is coordinate transformation matrix from Ss to S2, \u0424s is parameters of face gear tooth surface . Then equations of orthogonaln face gear tooth surface can be expressed by using two parameters \u03b8s and \u03c6s, it is:          \u00b1\u2212= += = )sin(cos ] cos cos )cos(sinsin [- ] cos sin -)cos(sin[cos bs2 2s 2 2bs2 2s 2 2bs2 \u03b8\u03b8 \u03b8 \u03b8\u03b8 \u03b8 \u03b8\u03b8 \u03d5\u03b8\u03d5 \u03d5 \u03d5 \u03d5\u03b8\u03d5\u03d5 \u03d5 \u03d5 \u03d5\u03b8\u03d5\u03d5 s s s rz q ry q rx \u2213 \u2213 (4) Where \u03c62=q2s\u03c6s When the symbol of \u03b8s, \u03c6s change, you can get another tooth surface of the face gear tooth. Modeling process of face gear and optimizing Simulating Face Gear Machining Fan campaign in the UG / Open GRIP is to simulate actual case of the cutting process. The revolution of tool and gear blank is realized with rotation transformation in grip. Gear cutting process is simulated by Boolean subtraction in UG. Then the model of the face gear is obtained. It is positive correlation between the precision of model and times of Boolean operation. According to the shaping principle [5, 6], a program can be built. As long as the times of Boolean operation between spur gear tool and face gear blank achieve a certain range, a precise face gear model can be got with this method. Considering the economical, the efficiency and precision at the same time, in the process of shaping a gear tooth, the number of Boolean operation is set as five in this article. However, in the initial model, the gear face is made up of a series of little face from Boolean operation. Regular meshes can\u2019t be got in the later finite element analysis. So the initial model must be optimized as follow. A file in stl format that can be used in Imaeware can be transform in UG. Then the tooth face of face gear can be obtained in Imaeware. The face with two cylindrical faces whose diameters are inner diameter and external diameter respectively and the two face in dedendum and addendum these face above form one tooth of face gear. Then array the tooth model and make Boolean operation, and a completed face gear model is finished that can be used for the later finite element analysis. Figure 2(a) shows the model before optimization, and (b) shows the optimized model. (a) model before optimization (b) model after optimization Fig.2"
        ]
    },
    {
        "image_filename": "designv11_100_0001996_1.5062329-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001996_1.5062329-Figure5-1.png",
        "caption": "Figure 5 An array of microneedles - insert (left) and moulded component (right)",
        "texts": [
            " This study should provide sufficient information about the technical requirements in designing and optimising the component technologies in the targeted manufacturing process chain for serial production of hollow microneedles\u2019 patches. The insert that was designed for this initial study had overall dimentions of 3mm x25mm x 35mm. The insert contains only seven needles so that the moulded microneedle patch could be assembled to a pre-existing experimental set-up for testing purposes. The location of the array of needles on the insert is given in Figure 5. To optimise the design tests were performed on a needle array component produced using an EnvisionTec rapid prototyping (RP) machine. This design study showed that the required holes can be successfully drilled by ps laser machining as depicted in Figure 6. Ns laser drilling would have been a faster and a more cost effective option but had limitations in terms of tolerances and minimum feature size. However, it should be noted that depending on the laser\u2019s wavelength, the polymer material might show no response to a ns laser processing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.4-1.png",
        "caption": "Figure 2.4. The definition of angles.",
        "texts": [
            ", referred to Sb; and Ca is the accelerating torque, in p.u. From (2.4) it results 2H d dt vr v0 \u00bc Ca \u00bc Cm Ce Pm Pe (2.400) In addition, denoting by v \u00bc vr v0 v0 \u00bc vr v0 1 (2.5) and taking into account that dv dt \u00bc d dt vr v0 (2.50) from equation (2.400) it results the equation of motion in per unit 2H dv dt \u00bc Cm Ce Pm Pe (2.6) Denoting by d the angle (in electric radians) giving the position of rotor, at an instant t, with respect to a synchronously rotating reference system (with v0), and by d0 its value at the instant t\u00bc 0 (see Figure 2.4a and b), then d \u00bc vrt \u00fe d0 v0t \u00bc vr v0\u00f0 \u00det \u00fe d0 resulting that dd dt \u00bc vr v0 \u00bc v0 vr v0 v0 \u00bc v0v (2.7) and d2d dt2 \u00bc v0 dv dt \u00bc v0 2H Cm Ce\u00f0 \u00de and thus we obtain another form of the equation of motion 2H v0 d2d dt2 \u00bc Cm Ce Pm Pe (2.8) Usually, the differential equation of motion contains also a damping torque component, obtained by adding a term proportional to the speed deviation v, in equations (2.6) and (2.8), respectively, 2H dv dt \u00fe Dv \u00bc Cm Ce Pm Pe (2.9) 2H v0 d2d dt2 \u00fe D v0 dd dt \u00bc Cm Ce Pm Pe (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure27-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure27-1.png",
        "caption": "Fig. 27 Heel chamfering with EM\u2014spiral bevel pinion",
        "texts": [
            " For one, when chamfering the Toe of a pinion with a small pitch cone angle, again the turn table angle is likely to exceed the machine\u2019s limit and collide with the tool spindle (left, Fig. 26). Beyond this, even a small diameter EndMill is likely to interfere with the concave tooth flank when chamfering the bottom of the tooth (right, Fig. 26) and therefore either the Pivot Angle must be reduced, or else this solution becomes unacceptable. Tool spindle to turn table collision is not likely to occur at Heel (left, Fig. 27); and if the Pivot Angle is correctly chosen, tool interference with the tooth flank can be avoided (right, Fig. 27). The Ball Mill tool (BM), thanks to its spherical end, can be fitted in places where an End Mill tool would not do an acceptable job. Consider for example the spiral bevel pinion shown in Fig. 28, left. The Ball Mill tool can be plunged vertically along the Toe and Heel edges without any risk of tool spindle to turn table interference. And by carefully selecting the Ball Mill diameter, the fillet area can also be chamfered. Five unit vectors (left, Fig. 29) are required to control the BM at any point along a tooth edge: Vo: N X T ,\u2212\u2212\u2192 Tool: the tool vector,\u2212\u2212\u2212\u2192 Trans: axis about which vector \u2212\u2212\u2192 Tool is rotated by the Pivot Angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002963_s0005117913080110-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002963_s0005117913080110-Figure1-1.png",
        "caption": "Fig. 1. Tractive vehicle and trailer.",
        "texts": [
            " (9) As compared with the control proposed in Theorem 1, in the general case this control has a rather complicated structure and uses the value of the last coordinate at the current instant, and the delay \u03c42 must be the same in all addends. On the other hand, this method allows one to determine an explicit, though only sufficient, estimate of the maximal delay. In virtue of arbitrariness of the positive control parameters a and b, at that the variables \u03c41 and \u03c42 can be made arbitrarily great, decrease in a and b and increase in \u03c41 and \u03c42 leading only to a lower rate of deviation convergence. To illustrate the above results, we consider the problem of controlling motion of a tractive vehicle with trailer (see Fig. 1). With the notation shown in the figure, the kinematic motion equations (see, for example, [16]) are given by x\u0307c = v cos \u03b80, y\u0307c = v sin \u03b80, \u03d5\u0307 = \u03c9, \u03b8\u03070 = v l tan\u03d5, \u03b8\u03071 = v d1 sin(\u03b80 \u2212 \u03b81). (10) We take v and \u03c9 as the control actions. For n = 5 (see [4]), the local change of variables can rearrange system (10) in (1). Let us consider by way of example the problem of motion stabilization along the straight line, which corresponds to (xrc yrc \u03d5r \u03b8r0 \u03b8r1 vr \u03c9r) = (t 0 0 0 0 1 0), or, in terms of the new variables, (xr1 xr2 xr3 xr4 xr5 ur1 ur2) = (t 0 0 0 0 1 0)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001340_b978-081551497-8.50005-6-Figure3.25-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001340_b978-081551497-8.50005-6-Figure3.25-1.png",
        "caption": "Figure 3.25 Boundary element matrix equations are composed mainly of boundary integrals if no interior source terms are present. Two special functions are needed to build the equations: the fundamental solution of the PDE \u03c8* and its associated flux q*, both depicted in the lower part of the figure for the Laplace equation. The symbols h and g represent the final matrices, and \u03c8k and qk are the unknown potential and surface flux density nodal values, respectively.",
        "texts": [
            "com programming C H 0 3 9 / 9 / 0 5 8 : 4 9 A M P a g e 1 3 2 The most distinguishing feature of the boundary element method (BEM)[29] is that the mesh is of one dimension lower than the computational geometry, i.e., only the boundaries between domains/materials are meshed. Since meshing is generally a tough job, this represents one of the advantages of the boundary element method. Typically, the same element polytopes are used as for the finite element method (see Fig. 3.24). The BEM can be used for the discretization of a PDE if the following conditions are met: \u2022 The PDE can be rewritten as a boundary integral (see Fig. 3.25). If volume integrals remain (as is the case, for example, for the Poisson equation), these can also be handled. The BEM loses one of its important advantages, however, because in this case a volume-filling mesh is also needed to perform the numerical integration. MEMS AND NEMS SIMULATION, KORVINK, RUDNYI, GREINER, LIU 133 CH03 9/9/05 8:49 AM Page 133 \u2022 There is a Green function for the PDE. In engineering terms, the Green function of a PDE is the analytical solution to a unit excitation. For example, in electrostatics, which is governed by a Poisson equation, the Green function is the potential field generated by a unit charge placed in infinite, empty space (see Fig. 3.25). The boundary element method permits a straightforward interpretation. Since we know the analytical response to a unit excitation, we can use the linear superposition principle by expressing every excitation as a weighted sum of unit excitations. Then the response of the system is simply the weighted sum of the analytical Green functions. Continuing with this engineering analogy, we remain with the electrostatic equation. On a boundary segment, we can either specify a voltage or we can place some charges, but we cannot do both"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002653_amr.189-193.2037-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002653_amr.189-193.2037-Figure5-1.png",
        "caption": "Fig. 5 The simulation of the face-gear. Fig. 6 The result of simulation of the face-gear.",
        "texts": [
            " The face-gear surface is determined as the envelope to the family of head-cutter surfaces as follows: ),(),,( 11111 \u03b2\u03d5\u03b2 SrMMMSr cbtqtq= (5) Here ),(1 \u03b2Src is the vector function that determines the surface of the head-cutter in coordinate system Sb; i is the tilt angle of the head-cutter; R1 is the pitch radius of the gear; B\u2206 is the sliding base; H is the horizontal setting of the head-cutter; V is the vertical setting of the head-cutter; \u03b4 is the spiral angle of the gear; 1\u03d5 is the work spindle rotational angle and             \u2212 = 1000 0cossin0 0sincos0 0001 1 ii ii M bt (6)             \u2206\u2212 +\u2212 \u2212 = 1000 100 )coscos(010 sin001 11 1 1 B Rir rR M c c qt \u03d5\u03b4 \u03b4 (7)             \u2212 = 1000 0100 00cossin 00sincos 11 11 1 \u03d5\u03d5 \u03d5\u03d5 qM (8) The simulation model of face-gear has been developed based on the generating algorithm. The machining simulation has been performed for a face-gear of common design parameters represented in Table 1. Fig. 5 represents the simulation process of face-gear. Fig. 6 shows the result of simulation of gear. Based on the performed research, the following conclusions might be drawn: (1)A new geometry of face-gear with curvilinear shaped teeth and non-zero spiral angle is proposed. The geometry is based on generation of face-gear by a tilted head-cutter. (2)The machining mathematical model of face-gear is determined analytically based on the spatial kinematics. (3) The simulation model for face-gear is developed based on the generating algorithm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002406_2011-01-1689-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002406_2011-01-1689-Figure2-1.png",
        "caption": "Figure 2. Elastic wedge location in gas turbine base frame",
        "texts": [
            " Figure 3 shows the results of a frequency analysis of the structure without elastic wedge at three different frequencies. The analysis shows that the larger displacements occur at the target beam indicated in that figure. In order to set the parameters of the elastic wedge the following process is followed. First, a single beam with large displacements, target beam in figure 3, is selected from the frequency analysis results of the structure. Once the target beam is selected, the connection point to the elastic wedge can be determined, figure 2. Next, the approximate diameter of the elastic wedge at the thicker edge is estimated by matching the impedance on the elastic wedge at the connection point with that of the target beam at the same point. Impedance matching ensures that energy will be transmitted from the target beam to the elastic wedge with little reflection at the connection point. The other parameters of the elastic wedge, length, diameter of thinner edge and power constant, are set based on experience, space available in the structure and manufacturing limitations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002691_imece2012-87321-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002691_imece2012-87321-Figure10-1.png",
        "caption": "Figure 10. HYDRAULIC SUSPENSION MODEL",
        "texts": [],
        "surrounding_texts": [
            "The main components of the wind turbine drive train are the main shaft on its bearing(s), the gearbox, the generator and the nacelle. All these components are included in the multibody model and modeled in a similar way. Therefore the main model features of the flexible multibody model are the bearings, gears, structural components, gearbox bushings and generator coupling. The following paragraphs discuss these in detail. Bearings in the gearbox and main bearing In the multibody models used in this paper the bearings are represented in a discrete fashion by a spring damper relationship. Diagonal 6x6 stiffness and damping matrices are used, which results in the following equations to represent the bearings:[ FBody,1 FBody,2 ] = KBearing. [ qBody,1 qBody,2 ] +CBearing. [ q\u0307Body,1 q\u0307Body,2 ] (1) with: Kbearing =  kax,ax 0 0 0 0 0 0 krad1,rad1 0 0 0 0 0 0 krad2,rad2 0 0 0 0 0 0 0 0 0 0 0 0 0 ktilt1,tilt1 0 0 0 0 0 0 ktilt2,tilt2  (2) Gears In order to minimize the influence of the specific design of the GRC gearbox on the generality of the overall drive train behavior, it was chosen not to include the gear meshing stiffness in the models. In theory the forces and moments which are introduced in the gearbox should be transferred to the gearbox bushings through a path comprising of the planet carrier, planet carrier bearings and the gearbox housing, as shown in Figure 11. Unless there is play in the bearings or the planet carrier 5 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76595/ on 03/22/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use and/or housing stiffness is insufficient the gear meshing stiffness does not play a role in this mechanism. In addition by excluding the gear meshing stiffnesses it is possible to assure that the housing is an important part of the transfer path. Since the gear meshing stiffness is needed to counteract the torque applied at the rotor, an equivalent total gear meshing stiffness is taken into account in the torsional DOF and superimposed on the stiffness values at both planet carrier bearings. Structural components All structural components were considered to be rigid, except for the planet carrier and gearbox housing. These components are represented by means of Craig Bampton reduced finite element models [6]. It was chosen not to consider the nacelle and main shaft as flexible to stay as independent as possible from the detailed design of the NREL GRC turbine. Since the flexible behavior is determined by the local details of the geometry of the component also the system response will be influenced by these local design choices. In case rigid models are used the forces acting on the different subcomponents of the turbine can be assessed in a more general way, which fits better in the scope of this paper. Gearbox bushings Linear 6x6 stiffness and damping matrices are used to represent gearbox bushings. Only the diagonal terms are considered. Stiffness values were delivered by the GRC. Reliability and vibration sensitivity assessment The first part of the analyses focusses on the investigation of the influence of the drive train configuration of the GRC turbine on the overall parameters important for the reliability of the gearbox. The results shown are specific for the GRC turbine. Nonetheless some general trends also valid for similar design layouts can be deducted from the performed analyses. For brevity only the results of the FRF analyses to the rear planet carrier bearing are discussed. Torque loading The main loading in the wind turbine drive train is torque loading. Due to the fact that the dynamic behavior of the drive train is coupled in the different DOFs this torque loading will also influence non torque DOFs at the rear 6 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76595/ on 03/22/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use planet carrier bearing and gearbox mounting bushings. Figures 12 and 13 show the responses in the radial DOFs for the three system layouts. It is clear that the three point mounting shows the highes amplitudes, especially at resonance. The hydraulic system creates the lowest response amplitudes in the planet carrier bearing. In addition to the higher amplitudes also a higher frequency is found for the first system resonance in the TPM. The frequency of this resonance peak is an important parameter in the overall turbine design, since it should be avoided to permanently excite this peak. Figures 14 and 15 show the transfer paths to the rotational displacements in the rear planet carrier bearing. Again the TPM results in the highest response amplitudes, whereas the lowest resonance frequency is related to the hydraulic system. Especially in the Ry DOF the hydraulic system decreases the response significantly. The freedom of the gearbox to move up and down in Z-direction is an important contributor to this reduced moment loading of the planet carrier bearings. The gearbox response to this moment loading is significantly determined by the play in the planet carrier bearings. Achieving minimal play in these bearings sets extra constraints on the overall gearbox design. Therefore the fact that the moment loading is minimized, is an advantage of the HS. Non-torque loading: Thrust loading In addition to the torque loading the rotor applies axial or thrust loads on the drive train. These loads should be isolated from the gearbox in order to minimize axial loads within the gearbox bearings. Higher axial loads in the gearbox require specific bearing configurations and are therefore an extra point of care for the design process. Figures 16, 17, 18 and 19. Similar to the previous section the TPM shows a significant resonance in the frequency range of interest and has the overall highest response amplitudes. From Figure 17 it is clear that the freedom of the HS in vertical direction removes the resonance of the suspension system from the frequency range of interest. With regard to moment loading due to the trust forces Figures 18 and 19 show that the HS system results in the lowest PC bearing torque loads about the Y- and Zaxis. The response in axial direction was not discussed since the planet carrier bearings have no axial stiffness and therefore all axial loading is diverted to the main bearing. 7 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76595/ on 03/22/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Non-torque loading: Ty loading Non-torque loading is an important aspect in gearbox reliability. In addition to the response at the planet carrier bearings in non-torque DOFs the moment loading at the rotor input is an important aspect. Figures 20, 21, 22 and 23 illustrate the respective transfer paths. For DOFs Y and Z similar conclusions can be drawn as in previous sections. With regard to Ry and Rz response in the planet carrier bearing it is clear that the HS scores better than the TBC and TPM, especially for the response in the Ry DOF. This is expected to be mainly due to the free suspension in vertical direction, allowing the gearbox to compensate for a significant part of the loading about Ry. Comparing the results for the TBC and HS in Figures 21 and 22 shows the added value of the free gearbox movement to reduce loads in the first planetary stage of the gearbox. Non-torque loading: Rz loading In addition to a nontorque loading at the rotor connection in Ry DOF the loading in Rz DOF was investigated. Figures 24, 25, 26 and 27 show the resulting responses at the rear planet carrier bearing; In general similar conclusions can be drawn as in previous paragraphs. In conlusion it can be stated that with regard to reliability both the TBC and HS show better behavior than the TPM for the GRC drive train layout. In the responses in the RY DOF the HS can additionally reduce the amplitudes of the rear planet carrier bearing responses. Moreover one of the resonance peaks in the Z response disappears from the frequency range of interest. However using the HS results in higher response amplitudes in other DOFs. Therefore based on the current study no general conclusions can be drawn about choosing either the TBC or HS. Nonetheless it can be stated that both the TBC and HS show significant advantages over the TPM. It is important to re- 8 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76595/ on 03/22/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use mark however that the TPM results can be improved by stiffening the planet carrier, planet carrier bearings and housing in order to make the system less susceptible to non-torque loading. Therefore accurate modeling during the design phase is required to assess the different loads in the different components of the drive train, especially to design the main shaft bearings and hydraulic stiffness of the HS."
        ]
    },
    {
        "image_filename": "designv11_100_0003848_imece2012-87624-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003848_imece2012-87624-Figure3-1.png",
        "caption": "FIGURE 3. APPROXIMATE KAI MODEL SCHEME IN THREE DIFFERENT MANIPULATOR POSES",
        "texts": [
            " In [13] an analytic approach to model the dynamics of a fire rescue turntable ladder is presented, where the flexibilities of the arm are approximated by the solution to an Euler Bernoulli Beam of variable length with a fixed tip mass. The rotational dynamics of the KAI manipulator extend this problem to different configurations of the manipulator in the vertical plane, influencing the horizontal eigenfrequencies and eigenvectors of the manipulator. The KAI Manipulator is modeled within the FMBS framework of [6] to capture the horizontal dynamics of the manipulator and to account for the influence of the vertical pose of the manipulator as depicted in Figure 3. The modeling of a flexible link element Ki is revisited. As well linear gain matrices, describing the velocity of a destination body in dependence of the velocities of a source body and the velocities of a rigid joint is derived, which are used to evaluate the jacobian of an open kinematic chain at a given deformation of the links and joint state.. The jacobian can be used to derive a minimal coordinate representation of the open kinematic chain. Approximate parameters of the KAI manipulator are given and the flexible multi body system model of the KAI manipulator is described",
            " xII n ]T denotes the global vector of the reference frame velocities of all bodies of the open kinematic chain. Replacement of xII with (11) and multiplication with JT the Euler Newton model in (3) can be written as JTMJ\ufe38 \ufe37\ufe37 \ufe38 :=M\u0302 \u02d9yII + JMJ\u0307yII = JThe + JThQ. (12) By neglecting the gyroscopic and centrifugal forces JMJ\u0307yII and using D\u2019Alemberts Principle JThQ = 0, it becomes M\u0302 \u02d9yII = K\u0302yI + D\u0302yII (13) with y\u0307I = yII . The KAI manipulator (Figure 1, Figure 2) can be modeled by an open kinematic chain, which is illustrated in Figure 3. Starting out from a turntable K0, the actual elastic manipulator consists of three flexible beam-like elements K1, K2 and K3 and the stiff tool elementK4. The Denavit-Hartenberg-Parameters of the stiff approximate to manipulator used within this work are TABLE 1. DENAVIT-HARTENBERG-PARAMETER OF APPROX- d a \u03b1 \u03c6 K0 0 0 \u03b10(t) \u03c0/2 K1 0 5.0 \u03b11(t) 0 K2 0 4.0 \u03b12(t) 0 K3 0 3.0 \u03b13(t) 0 K4 0 1.0 0 \u2212\u03c0/2 TABLE 2. GEOMETRIC PARAMETERS OF FLEXIBLE LINK ELEMENTS l IT [ cm4 ] Izz [ cm4 ] Iyy [ cm4 ] Ax [ m2 ] K1 5",
            " (17) The turntable of the manipulator is driven by a hydraulic system of high internal stiffness, therefore the dynamics of the turntable \u03b10 can be neglected. Within the topic of this paper, the turntable is assumed to be clamped \u03b1\u03070 = 0, which can be included into (17) by model reduction techniques. As shown in Figure 6, the first eigenmode of the KAI manipulator represents a similar bending of the flexible elements in the horizontal plane. However the composition of shape functions within each eigenmode, as well as the corresponding eigenfrequency depends on the pose of the manipulator. The tree different exemplary poses are shown in Figure 3. In Table 4 the eigenfrequency of the first mode is listed at each exemplary pose. The shape function coefficients of the first two shape functions of each flexible link element are shown, qi1 corresponds to the the shape function \u03a6i1, which is the torsion mode of the body Ki and qi2 denotes the coefficient of the first horizontal bending mode of Ki. The frequency of the first mode is highest for the fully extended arm in Pose C. In pose C, the frequency gets up to approximately 0.7Hz, as well the coefficients of the shape functions vary in dependence of the pose"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002407_amr.189-193.1409-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002407_amr.189-193.1409-Figure1-1.png",
        "caption": "Fig. 1. The RRR limb with constraint 1 2 3, ,f f f",
        "texts": [
            " Thus the feasible limbs of spherical parallel mechanism could be divided into four types: the 0-constraint limb, the single-constraint limb, the two-constraint limb and the three-constraint limb. The following chapter will introduce the synthesis procedure of the limb for spherical parallel mechanism. Since the limb constraint system consist three constraint forces 1 2 3, ,f f f , the twists of limb can be obtained: 1 1 1 1 2 2 2 2 3 3 3 3 [ , , ; 0, 0, 0] [ , , ; 0, 0, 0] [ , , ; 0, 0, 0] R a b c R a b c R a b c = = = Where 1 2 3, ,R R R denote the revolute joints whose axis passing across the origin. Thus the limb with three-constraint is the RRR chain shown in Fig. 1. Since the limb constraint system consist one constraint force 1f , the twists of limb can be obtained: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= , 1 4 4 4[0,0,0; , , ]P d e f= , 2 5 5 5[0,0,0; , , ]P d e f= The new revolute joints 4 5,R R can be obtained by the linear combination of revolute joint and prismatic joint, so the twist system of limb could become: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= 4 1 1 1 4 4 4[ , , ; , , ]R a b c d e f= , 5 2 2 2 5 5 5[ , , ; , , ]R a b c d e f= Therefore, the UPU limb can be obtained to afford the constraint force 1f , as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003559_amm.268-270.1063-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003559_amm.268-270.1063-Figure1-1.png",
        "caption": "Figure 1. Coordinate system of face gears",
        "texts": [
            " They are small size of transmission system, light weight, less vibration and noise and a wide range of applications in the aerospace industry. Because of the gap and errors of face gear the separation is often happened. This phenomena attract the attention of many researchers. But many research focus on the frequency analysis[5]-[6]. But the system is a complex system. Above all the system is nonlinear and unstable. So the method is not effectual. In the paper a model is established. It include four case. It can simulate the real transmission situation. Mathematical model of the orthogonal Face Gear Figure 1 shows the orthogonal coordinate system in face gear machining process, the angle of face gear and the cutting tool is 90 \u00b0.The cutting tool involute tooth surface equation in the coordinate 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-27/05/15,16:24:51) system is [5]:             +++\u2212 +\u2212+\u00b1 = 1 ]sin()[cos( )]cos()[sin( ),( 00 00 s ssksssbs sssssbs sss u r r ur \u03b8\u03b8\u03b8\u03b8\u03b8 \u03b8\u03b8\u03b8\u03b8\u03b8 \u03b8 (1) Where rbs is the cutting tool involute radius of base circle; su is parameters of cutter tooth surface point on the axial direction; s\u03b8 is angle parameters of point in cutter involute line; s0\u03b8 is angle from cutter alveolar symmetrical line to the involute starting point), \"\u00b1\"symbol correspond to the both sides of cutter alveolar involutes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure48-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure48-1.png",
        "caption": "Fig. 48 CoSIMT pivot to avoid interference\u2014Toe/Heel chamfering",
        "texts": [
            " Gosselin \u2212\u2192 Vo: N X T ,\u2212\u2212\u2192 Tool: vector parallel to the tool axis,\u2212\u2212\u2212\u2192 Pivot : axis about which vector \u2212\u2212\u2192 Tool is rotated by the desired chamfer angle. \u2212\u2192 Vo is obtained from the cross product of N and T . To avoid working on the blade tips, the CoSIMT slides along vector T by distance Ds; this is therefore a \u201cMoving Contact Point\u201d that ensures evenwear along the cutting edge of the tool and improves tool life. In the case where the CoSIMT OD is too large, the tool can be pivoted out of the mesh around axis \u2212\u2212\u2192 Tool, as shown in Fig. 48. Finally, by carefully selecting the CoSIMT tool dimensions, it is also possible to target the root area of the Toe and Heel edges. Fig. 49 shows\u2014in several different positions\u2014a CoSIMT chamfering a straight bevel pinion tooth Toe edge: it is clear that the fillet and root areas of the tooth gap can be targeted by carefully selecting the CoSIMT dimensions, Chamfer Angle and Pivot Angle. The sameapproach is used for spur, helical and spiral bevel gears andwill therefore be omitted here. The intent of this chapter was to explore some of the deburring and chamfering options that are offered to gear manufacturers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002774_mmar.2012.6347894-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002774_mmar.2012.6347894-Figure1-1.png",
        "caption": "Fig. 1. Basic structure of robot motion",
        "texts": [
            " The paper proposes a new method for passing close to hand-wrist singularities, in which the errors are concentrated in prescribed task space directions, thus avoiding uncontrolled end-effector motions which could lead to collisions with the fixed target. An industrial robot with kinematically consistent robot axis control is employed, so that the robot dynamics need not be considered. Such aspects will be tackled in future publications. We regard a robot whose base is moving with respect to the inertial system K0 as measured by an inertial measurement unit (IMU) with gyroscope-fixed frame Kg (Fig. 1). Displacements are measured with respect to an assumed average center of displacements in directions (i) heave zg, (ii) sway yg, and (iii) surge xg; orientation is measured in absolute angles in the sequence (i) heading \u03c8g, (ii) pitch \u03b8g, and (iii) roll \u03d5g in accordance with the standard notation for motion description (SNAME [2]). Collecting all IMU values in a vector \u03d5g = [xg , yg , zg , \u03c8g , \u03b8g , \u03d5g] T , (1) the pose of the IMU frame Kg with respect to the reference frame K0 is thus described by the homogeneous transformation 0Ag = [ 0Rg(\u03d5g) g 0rg(\u03d5g) 0 1 ] , (2) with 0Rg(\u03d5g) = Rot [ z, \u03c8g ] \u00b7 Rot [ y, \u03b8g ] \u00b7 Rot [x, \u03d5g ] 0 0rg(\u03d5g) = Trans[x, xg ] \u00b7 Trans[ y, yg ] \u00b7 Trans[ z, zg ] , where a brc denotes a radius vector from the origin of frame Kb to the origin of frame Kc in coordinates of frame Ka"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002683_amr.443-444.282-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002683_amr.443-444.282-Figure5-1.png",
        "caption": "Figure 5. The underwater vehicle three-dimensional modeling.",
        "texts": [],
        "surrounding_texts": [
            "PID controller in this system is to control underwater vehicle noumenon, and take error E and error change rate EC as the inputs of fuzzy controller, according to PID fuzzy reasoning method the output of parameters are KP, KI and KD of variation for online tuning. To meet the different error E and error change rate EC requirements on the different parameters of the controller so that the controlled object has a good dynamic and static performance. The control equation for the PID controller\n0\n1 ( ) ( ) [ ( ) ( ) ] dE t u t KP E t E t dt KD\nKI dt\n\u03c4 = + +\u222bi\n(7)\nE(t): error. u(t): The output control for the controller. The fuzzy controller KP, KI, KD rules based on E, Ec is the core content and specific as follows:\nWhen\uff5cE\uff5cis larger, to take a larger KP and a smaller KD for a good tracking performance, in\norder to avoid the system appearing the larger overshoot, it usually take KI=0;\nWhen\uff5cE\uff5cand\uff5cEc\uff5care medium, for the system with a small overshoot, to make KP small. In\nthis case, the KD value influenced the system, should be small. And the KI value should be\nappropriate;\nwhen\uff5cE\uff5cis small, to get a good stability, KP and KI should take a little bigger, considering to\navoid Oscillation occurs when the system settings, and the system anti-interference performance;\nWhen \uff5cEc\uff5cis larger, and the smaller KD is desirable;\nWhen\uff5cEc\uff5cis smaller, the desirable KD could be a little larger.\nThe input and output variables of the fuzzy language to describe,Set the input variable\uff5cE\uff5c\nand \uff5cEc\uff5clinguistic fuzzy subset is (negative large, negative, negative small, zero, small, middle, large, and abbreviated as (NB, NM, NS, ZE, PS, PM , PB), the error E and error change rate of Ec volume Of the (-3,3) of the region. Similarly, design the output Kp, Ki Kd of fuzzy sets as (NB, NM, NS, ZE, PS, PM, PB), and quantified to the region (-3, 3) inside. The input and output variables of the membership function curve respectively, as shown",
            "This paper obtains the underwater robot navigation control simulation, other simulation results can be obtained through Matlab, and the results are similar. System simulation diagram shown in figure 4 can be seen, with the Fuzzy PID controller effect, the controller of the existence of a faster speed to close the initial steady state, when near the steady state convergence rate slows down, presents almost no overshoot in order to meet the desired purpose.\nBased on the speed, heading and depth control to validate the strategy of the feasibility of the AUV control, its tail is equipped with axial thrusters, voltage by controlling the four DC brushless motors to control its forward speed, heading and depth.\nMore than the Fuzzy PID controller use genetic algorithm optimization for the fuzzy control rules and parameters calculating overshoot, to improve the response speed of underwater vehicle\u2019s control system, low overshoot and characteristics of static error. At the same time in the system design optimization, we will use Lyapunov stability theory to analysis the stability on underwater currents of shock response model under the stability analysis research. Through the study, it will solve the middle-controlled key issues of the anti-surge control design, optimization and stability.\nThis paper has analyzed the control problems of the surge in shallow waters and designed specific control systems for underwater vehicles in shallow waters.\nThe fuzzy PID controller is designed to achieve the intended purpose of overshoot suppression feature, the further work is to improve the a comprehensive strategy based on fuzzy PID controller including the early prediction and management of late, under the impact of the surge speed, stability, Effectively achieve the intended purpose. The application provides a theoretical basis and new ideas and technology for underwater vehicles\u2019 application in shallow waters.",
            "The authors would like to acknowledge Shanghai Education Commission Foundation for providing financial support for this work under grant No. 09zz091.\n[1] LCDR Jeffery S. Riedel, Healey, Anthony J. Shallow water Station Keeping of AUVs Using\nMulti-sensor Fusion for Wave Disturbance Prediction and Compensation. Oceans Conference\nRecord (IEEE), Vol. 2, 1998:1064-1068\n[2] Li Shuyong, Wang Danwei, Poh Engkee. Nonlinear Adaptive Observer Design for Tracking\nControl of AUVS in Wave Disturbance Condition OCEANS 2006-Asia Pacific, 2007: 1-8\n[3] Liu Shuyong, Wang Danwei, Poh Engkee, Chia chin swee. Nonlinear Output Feedback\nController Design for Tracking Control of ODIN in Wave Disturbance Condition Proceedings\nof MTS/IEEE OCEANS, Vol.2, 2005: 1803-1810)\n[4] Poorya Haghi, Mahyar Naraghi, and Seyyed Ali Sadough Vanini. Adaptive Position and\nAttitude Tracking of an AUV in the Presence ofOcean Current Disturbances. 16th IEEE\nInternational Conference on Control Applications Part of IEEE Multi-conference on Systems\nand Control.Vol.1 2007:741-746\n[5] Silvia M Zanoli, Giuseppe Conte.Rometely operated vehicle depth control, Control\nEngineering Practice, Vol.11,2003:453-459\n[6] Wang ZhiXue, Bian Xinqian, Liu YunXia, Wang KuiMin. Near surface water power of AUV\nmotion control, Marine engineering research.Vol.28 No.5, 2006:63-66"
        ]
    },
    {
        "image_filename": "designv11_100_0003419_amr.199-200.449-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003419_amr.199-200.449-Figure8-1.png",
        "caption": "Fig. 8 The allocation of the Hall sensors in the static oblique coordinates",
        "texts": [
            " The magnetic pole of the magnetized ball is laid perpendicularly to the rotational axle by a vacuum pump. The axle rotational speed is 1800rpm. The variation of the magnetic pole caused by the rotation of the magnetized ball is detected by a Hall sensor which is attached on a precision rotational plate. The Hall sensor, HW-300A, is produced by Xuhuacheng Electric Company with an input voltage of 1.5V. The output voltage is amplified 100 times by a amplifier, transformed by an A/D transformer and acquired by a computer. Figure 8 shows the allocation of the Hall sensors in the static oblique coordinates. A point H(XH, YH, ZH)=(R, 0, 0) on axis X represents a Hall sensor which can be rotated by a precision rotational plate. When H is rotated about axis Z, the angle in XY plane is defined as \u03a8. When H is rotated about axis Y, the angle in XZ plane is defined as \u03b8. A point H\u2019(XH\u2019, YH\u2019, ZH\u2019) on axis X\u2019 is given as \u03b8coscos' \u03a8= RX H (15) \u03a8= sin' RYH (16) \u03b8sincos' \u03a8= RZ H (17) Figure 9 shows a test result of the Hall sensor. When \u03a8=60\u00b0 and \u03b8=10\u00b0, the voltage amplitude of H\u2019 is half of that of H (\u03a8=0\u00b0, \u03b8=0\u00b0)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure46-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure46-1.png",
        "caption": "Fig. 46 CoSIMT Toe chamfering both tooth flanks using the outside blade\u2014straight bevel gear",
        "texts": [
            " In addition, while chamfering the left tooth edge with the CoSiMT\u2019s outside blade\u2014i.e. the blade on the opposite side of the arbor\u2014causes no risk of interference between the tool\u2019s arbor and the work piece (left, Fig. 45), using the inside blade\u2014 the blade on the same side as the arbor\u2014requires an elongated arbor (right, Fig. 45) to avoid interference and this is likely to lead to vibrations and uneven results. This implies that, for practical reasons, the CoSIMT outside blade should be used to chamfer both the left and right tooth Toe and Heel edges, as shown in Fig. 46. This also implies that the process is not continuous: all the Toe/Heel edges on one tooth flank are done, and then all the edges of the other tooth flank. In practice, this does not involve excessive cycle time and is well worth given all the chamfering can be done using a single tool. Five unit vectors (left, Fig. 47) are defined to orient the CoSIMT at any point along the Toe/Heel edge of a spur or straight bevel gear: such that cutting blades work away from their tips; N : the local normal vector, 182 C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001815_978-3-319-02114-0_1-Figure1.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001815_978-3-319-02114-0_1-Figure1.1-1.png",
        "caption": "Fig. 1.1 Cyclic voltammogram of a Pt electrode contacting a 0.1 mol/L HClO4 ? 0.4 mol/ L NaClO4 ? 1 9 10-3 mol/ L o-aminophenol (o-AP) solution, pH 1. Scan rate, v = 0.1 V s-1; electrode area, A = 0.126 cm2 [1]",
        "texts": [
            "2 Electropolymerization of o-aminophenol in Acidic Media Barbero et al. [1] reported a study about the oxidation of orthoaminophenol (o-AP) and closely related compounds employing electrochemical, chemical, and spectroscopic measurements. The electro-oxidation of o-AP was studied on different electrode materials (Pt, Au, and glassy carbon (GC)) and different electrolyte media (1 \\ pH \\ 7). A typical voltammogram of a Pt electrode in contact with a 0.1 mol/L HClO4 ? 0.4 mol/L NaClO4 ? 1 9 10-3 mol/L o-AP aqueous solution (pH 1) is shown in Fig. 1.1. On the first positive sweep two peaks are defined (a) at 0.60 V (SCE) attributed to the oxidation of o-AP to the monocation radical (o-AP\u2022+) and another peak (b) at 0.85 V, which was assigned to the oxidation of (o-AP\u2022+) to dication. On the negative sweep none of these peaks show complementary peaks, indicating chemical follow-up reactions giving products detected as peaks c\u2013c0 and d\u2013d0 on the subsequent sweeps. It was observed that the system c\u2013c0 diminishes after continuous cycling in the same way as a, but the peak system d\u2013d0 increases, also showing the characteristic behaviour of a deposited electroactive substance",
            " Barbero et al. [1] also prepared the electroactive polymer by chemical oxidation of o-AP, and its properties were compared with those of the electrochemically produced POAP. The chemical synthesis of POAP confirmed that the actual monomer in the formation of the polymer is the cyclic dimer of o-AP, 3APZ. A calculation of a global value of the rate constant for the dimerization reaction was carried out in Ref. [1]. With regard to the solution pH, a similar voltammetric response to the one shown in Fig. 1.1 was observed for pH values lower than 3. At pH values between 3 and 7, a diminished current of the peak system d\u2013d0, assigned to the film formation, was observed. Barbero et al. [1] remark that if extreme care is not taken in the preparation of a POAP film, not only in the concentration but also in the potential ranges, the possibility of side reactions and consequently \u2018\u2018side\u2019\u2019 polymers, increases, and the real structures of the films obtained could be quite complex. Then, the best conditions proposed in Ref"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.35-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.35-1.png",
        "caption": "Figure 10.35 Arrangements of magnets of line-start IPM motors: 1, stator and rotor iron laminations; 2, permanent magnets; 3, non-magnetic material; 4, stator slots; and 5, rotor slots. Magnetic bridges are part of the rotor laminations to maintain the integrity of rotor laminations",
        "texts": [
            "80) where lf e is the length of magnet along the shaft direction and usually equal to the rotor lamination stack length. At the operating temperature, the above parameters are replaced by their respective values. It is possible that the demagnetizing curve becomes nonlinear at the operating temperature. In this case, Fceq should be used in place of Fc as shown in Figure 10.33. The Norton equivalent of a cuboid magnet is shown in Figure 10.34. A generic, circumferentially magnetized IPM rotor configuration is shown in Figure 10.35. The equivalent magnetic circuit of this configuration is shown in Figure 10.36, where R\u03b4, Ry1, Ry2, Rt1, Rt2, R\u03c3 , R1, R2, RS are the reluctances of the air gap, stator yoke, rotor yoke, stator teeth, rotor teeth, assembly gap between magnets and laminations, magnetic bridge I, magnetic bridges II and III (combined due to symmetry), and leakage through the rotor slots and the non-magnetic material respectively. End effects are neglected. Fluxes passing through these bridges are leakage fluxes. These magnetic bridges are highly saturated as can be seen from Figure 10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002528_s0219455412500393-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002528_s0219455412500393-Figure11-1.png",
        "caption": "Fig. 11. Example 1: Model problem de\u00afnitions.",
        "texts": [
            "10), Jp and Jt are the polar moment and the fourth moment of the A area about the shear centre, respectively. Finally, the solution scheme is based on the stationary problem for the total functional obtained by the sum of \u00b0exural, torsional and axial contribute \u00bc f \u00fe t \u00fe at: \u00f06:11\u00de A predictor-corrector scheme as described in Lopez,21,22 for identifying the equilibrium path is used in the analysis. Now we show a classical benchmark to illustrate the results obtained using the described corotational approach and the new proposed approach. The narrow cantilever beam shown in Fig. 11 was analyzed. The numerical results obtained in Battini and Pacoste10 can be taken as reference ( ). The wc vertical 1250039-20 In t. J. S tr . S ta b. D yn . 2 01 3. 13 . D ow nl oa de d fr om w w w .w or ld sc ie nt if ic .c om by M C M A ST E R U N IV E R SI T Y o n 03 /1 0/ 13 . F or p er so na l u se o nl y. load parameter \u2014 lateral tip displacement curves have been computed for the proposed approach and for the corotational approach in the case of expansion series, in Rodrigues formula, stopped at the second and third order"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003335_mhs.2013.6710438-Figure2.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003335_mhs.2013.6710438-Figure2.1-1.png",
        "caption": "Fig 2.1 the work of \u201cgripping the bolt on the tool box\" in CG",
        "texts": [
            " There are problems in recognition of the bolt, such asfewfeature points, the simplicity of the shape,which is why we conduct recognition of extracting the color region of the bolt in this time. Following this introduction, we introduce the work with respect to \"grip the bolt of tool box\" in Section 2, describe how to recognize the bolt in Section 3, show the experimental results in Section 4, consider the results along the method proposed in Section 5, and give expression to the conclusions in Section 6. 2. Gripping the bolt on the tool box The following has been brieflydecidedthe work procedure of work \"gripping the bolt on the tool box\". Figure 2.1 shows the image in CG of the work of \"gripping the bolt on the tool box\" by the semi-autonomous power distribution line maintenance robot. Moving the robot arm up to the toolbox. Using hand-eye camera, and recognizing the bolt on tool box. Let Heart of the recognition frame of the bolt come to the center of the hand-eye camerawhile controlling the robot arm to approach the vicinity of the bolt. Align the bolt, to grip the bolt. Pull the bolt from the tool box. Pulling out the bolt from the tool box"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003082_icems.2013.6754391-Figure13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003082_icems.2013.6754391-Figure13-1.png",
        "caption": "Fig. 13. Distribution of magnetic flux density (2 poles\u21926 poles).",
        "texts": [],
        "surrounding_texts": [
            "Fig. 10 shows the core loss of the stator at no load. If the number of poles in the motor is reduced, the lower frequency in the motor leads to a decrease in its core loss. The motor can decrease the core loss in the high speed region by about 34.3%. Therefore, pole changing leads to high efficiency in the motor."
        ]
    },
    {
        "image_filename": "designv11_100_0003445_ht2012-58139-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003445_ht2012-58139-Figure6-1.png",
        "caption": "Fig. 6: Dimensions of the rectangular stainless steel thin plate (all dimensions are in mm).",
        "texts": [
            " Surfaces ADD\u2019A\u2019, ABB\u2019A\u2019 and DCC\u2019D\u2019: These three faces have no constraints defined and are hence free to deform under the thermal load. BCC\u2019B\u2019: It is assumed that the workpieces are lying on a horizontal table during welding. This is incorporated in the FE solver using the compression only boundary condition. Initial condition: As for the weld pool dynamics modeling, it is assumed that the workpiece is initially at a uniform temperature distribution, given by Tinit. For the structural (stress) analysis, it is assumed that the workpiece is initially stress-free. Figure 6 depicts the geometrical dimensions of one of the two thin stainless steel plates are welded together with a butt joint. The geometry was created in the ANSYS Design Modeler and imported into ANSYS Meshing, where a structured mesh was created. Since the problem is symmetric about the weld centerline, the symmetry boundary condition has been employed while developing the computational domain. The meshed three-dimensional geometry, representing the symmetric half (Fig. 6), comprised of 0.4 million hexahedral cells. In order to accurately capture the weld pool dynamics, the mesh nodes were clustered in the vicinity of the welding path. The meshed computational domain is not shown here. The commercially available flow solver, ANSYS FLUENT \u00ae , has been used for the present simulations of weld pool dynamics. It provides an excellent user interface that enables the user to include alternate models via user-defined functions (UDFs) when the built-in models lack the desired degree of realism and/or accuracy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002372_icsens.2012.6411252-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002372_icsens.2012.6411252-Figure1-1.png",
        "caption": "Figure 1. the design of (a) the disk shaped microelectrode and (b) the three-dimensional double coils microelectrode chip",
        "texts": [
            " Scanning electron microscope (SEM) analysis was carried out to determine the morphological structures of the modified gold electrode using a S-4800 field This work is supported by the National Basic Research Program of China (973 Program) (Project Number 2009CB320300), the \"Strategic Priority Research Program\" of the Chinese Academy of Sciences (Grant No. XDA0602010) and the State Hi-tech Research and Development Program (863 Program) of China (Grant No. 2012AA040506). ) 978-1-4577-1767-3/12/$26.00 \u00a92012 IEEE emission scanning electron microscope (FE-SEM) produced by Hitachi (Japan). Two kinds of microelectrode chips, three-dimensional double coils microelectrode chip (Fig. 1 (a)) and disk shaped microelectrode chip (Fig. 1 (b)), were designed in order to study the impact of electrode microstructure on the electrochemical performance. As shown from Fig. 1, two kinds of electrodes have the same area of working electrode and counter electrode, and the area of working electrode equal that of counter electrode. The whole area of the two kinds of electrode is designed both 1 mm2. The three-dimensional double coils microelectrode chip is based on a pair of coil electrodes with 23 \u03bcm width, 17 \u03bcm spacing, and 20\u03bcm heights. On the one of the pair electrodes, working electrode, gold nanoparticles were modified and pyruvate oxidase was immobilized by mixed SAMs technology (Fig. 1 (b)). fabrication This three-dimensional double coils microelectrode chip was fabricated by the following process: Firstly, after thin positive photoresist (AZ1500) was spincoated and patterned onto the glass wafer, the bonding pads of platinum (Pt) were deposited by magnetron sputtering and liftoff process. Then Au layer was deposited onto the wafer using magnetron sputtering method. Secondly, negative photoresist (NR2-8000P, Futurrex) was spin-coated onto the wafer. After exposing and developing, a master mode of the negative photoresist was formed for the three-dimensional double coil microelectrode chip"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003043_icinfa.2013.6720454-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003043_icinfa.2013.6720454-Figure3-1.png",
        "caption": "Fig. 3. Sampling analysis of squatting motion",
        "texts": [
            " As there is symmetry between the left and right legs during the squatting motion, only left leg is analyzed. Joint 1, 4 and 5 participate in the motion during the whole squatting process which is shown in Fig. 2. As depicted in Fig. 2, joint 1 and joint 5 are always on the same line. Joint 1 does rectilinear motion while joint 4 does curvilinear motion. Joint 1 moves in a straight line while joint 4 moves along a curved path. Sampling analysis is done for the whole squatting process and the result is shown in Fig. 3. In Fig. 3, squatting motion begins at time t0 and ends at tn. By analyzing the beginning state and end state of joint 1, following constraints can be got: 1 0 1 1 0 1 2 ( ) 0 ( ) 0 ( ) 0 ( ) n n st sq v t v t z t z t L =\u23a7 \u23aa =\u23aa \u23a8 =\u23aa \u23aa =\u23a9 (2) where v1 is the velocity of joint 1 at time ti, z1(ti) is the displacement of joint 1 in the direction of Z axis of the world coordinate system at time ti and Lst2sq is the displacement of joint 1 from the beginning to the end of the squatting motion. Similarly, there are constraints for joint 4 as follows: ( ) 4 0 4 4 0 4 5 ( ) 0 ( ) 0 ( ) 0 ( ) 1 n n crus v t v t z t z t L \u03b8 =\u23a7 \u23aa =\u23aa \u23a8 =\u23aa \u23aa = \u2212\u23a9 (3) where v4 is the velocity of joint 4, z4(ti) is the displacement of joint 4 in the direction of Z axis of the world coordinate system at time ti, Lcrus is the length of the robot\u2019s crus and \u03b85 is the rotation angle of joint 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003258_icems.2011.6073613-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003258_icems.2011.6073613-Figure1-1.png",
        "caption": "Fig. 1. Permanent magnet arrays. (a) periodic linear. (b) periodic planar. (c) non-periodic linear. (d) non-periodic planar.",
        "texts": [
            ". INTRODUCTION Permanent magnet arrays are widely used in some operating actuators such as linear motors and planar motors. These devices often consist of a periodic permanent magnet array as shown in Fig. 1 (a) (b) in order to obtain a large traveling range. A common method to analyze this type of permanent magnetic arrays is to solve Laplace equation under the assumption that the magnetic array is infinite [1]. However, this method is not accurate enough for some short-stroke precision actuators with non-periodic arrays as shown in Fig. 1 (c) (d) due to disregard of the real fringe field. In this paper, an analytical method based on the concept of magnetic charge and the method of image [2, 3] is employed for calculating the magnetic field of non-periodic arrays. II. ANALYTICAL MODEL A. Magnetic Field Due to a Parallel Magnetized PM According to equivalent magnetic charge method, the effect of magnetization for parallel uniform magnetized magnet can be represented by two magnetic charge surfaces on the sides of the permanent magnet, which are perpendicular to the magnetization direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003450_20110828-6-it-1002.00382-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003450_20110828-6-it-1002.00382-Figure5-1.png",
        "caption": "Fig. 5. Forces acted on the underwater vehicle.",
        "texts": [
            " Traditional yaw angle modeling algorithms are mostly based on hydrodynamic analysis, which may be not very suitable in nonideal fluid in practical applications. Therefore, system identification based on real-time sensing data has stronger application prospect. The swimming vehicle acts on the water with several main forces: gravity G , buoyancy force bF , water resistance in swimming direction hF , thrust by left long-fin leftF , thrust by right long-fin rightF and water resistance in section direction wF . Fig. 5 shows the forces acted on the vehicle in water. Here, wF is generated when leftF is not equal to rightF . Obviously, there are three factors having influence on yaw angle, which are leftF , rightF and wF . leftF and rightF are generated by the two long-fins and controlled by the oscillating frequency and oscillating amplitude of the longfin. And the direct influence on wF is the angular velocity through force analysis. For simplification, a stable value is chosen as the oscillating amplitude of the two long-fins"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003141_amm.86.908-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003141_amm.86.908-Figure4-1.png",
        "caption": "Fig. 4 Backlashes in radial direction, normal direction and circumstance direction",
        "texts": [
            " The data in Table 1 indicate that the maximum deformation appears on the eleventh point, and the x direction single side shrink is 0.0739mm. The minimum deformation is on the position of point 1 with the x direction single side shrink of 0.0135mm. To get the backlash change, the tooth root radial shrink should be transformed into normal backlash decrease. The radial backlash decrease of spline pair was similarly taken as the single radial shrink deformation. The relation of backlashes in radial direction, normal direction and circumstance direction [2] are shown in Fig. 4. In Fig. 4, jr is the radial backlash, jn is the normal backlash and jw is the circumstance backlash. And the pressure angle \u03b1 of the spline is 30\u00b0. So, the maximum normal backlash decrease is: \u25b3jnmax=2\u25b3jrmaxsin\u03b1=0.0739 mm And the minimum normal backlash decrease is: \u25b3jnmin=2\u25b3jrminsin\u03b1=0.0135 mm In order to avoid the conflict caused by the diminish of spline pair backlash, the normal backlash decrease should be amended according to the minimum normal backlash decrease \u25b3jnmin. At the same time, because of the different shrink quantity along the axial direction of spline pair, the outer spline tooth should be axial relieved to ensure the same normal backlash along axial direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003267_amr.314-316.1163-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003267_amr.314-316.1163-Figure5-1.png",
        "caption": "Fig. 5 Schematic diagram of tensile specimen sampling",
        "texts": [
            " 4, the maximum values of the micro-hardness in all measuring lines were in the fusion zone, and the smallest micro-hardness of the welded joint located in weld zone. Form Fig 4, the maximum micro-hardness each line is 220HV, 230HV and 280HV, and the micro-hardness constantly increase form top to bottom of weld joint. It can be seen that effect of welding thermal cycle on micro-hardness of welded joint is great. Tensile strength. Welding joint along the thickness direction was divided into Part A and Part B, and samples from these two parts are machined to measure the strength. Sample A is made from the upper zone and sample B is from the bottom zone. Fig 5 is the schematic of the sampling position. After the measurement and observation, the fractures appear in the welded seams, and the results of tensile can be seen in Table 1. welded joint. The tensile strength of the bottom of welding joint is 5.9% larger than that in the upper of welding joint. Microstructure and mechanical properties are closely linked. Multi-pass welding method makes the grains in welded joint at the bottom of the weld metal more uniform and finer, and the grains in the upper side of the weld metals remain the original columnar style which induces some acicular ferrite and Widmanstaten structure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003120_icate.2012.6403418-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003120_icate.2012.6403418-Figure7-1.png",
        "caption": "Fig. 7. Comparison between the flux density distribution lines for the motor of type M3AA 71 B2/4, 3000 rpm at t=0,5 s for winding : a) Double layer b) Simple layer",
        "texts": [],
        "surrounding_texts": [
            "The analysis is performed similarly for speed of 3000 rpm. In the first stage is comparing the distribution of flux lines in the motor at t = 0.5 s. The normal component of airgap magnetic flux density has the amplitude 1,2 T for motor with double layer winding and 1,4 T in the case of motor with single layer winding. Electromagnetic torque and speed variation is shown comparatively in the following figure: From the analysis of the above figure is observed that for both types of winding, the electromagnetic torque records pulsations which are found also in the motor speed to the stable function. In the case of motor with double layer winding pulsations amplitude is lower with 2 Nm than the pulsation torque amplitude of double layer winding motor. Magnetic pressure is calculated on the stator surface for each of the two types of windings. a) For double layer winding In the case of double layer motor windings of the amplitude average magnetic pressure is recorded all the time at the value of 0.94N. The most important harmonics orders are those presented in Table III. TABLE III. ORDER AND FREQUENCY OF MAJOR PEAKS FOR THE MOTOR OF TYPE M3AA 71 B2/4, 3000 RPM DOUBLE LAYER WINDING The order of harmonics Freqency [Hz] 11 550 12 600 14 700 18 900 26 1300 38 1900 39 1950 50 2500 62 3100 80 4000 81 4050 b) For single layer winding Fig. 11. The magnetic pressure exerted on the stator surface of the M3AA 71 B2/4 induction motor, single layer winding, 3000 rpm and amplitude of the first 90 harmonics In this case the amplitude average of the magnetic pressure has all the time the value of 0.94N. The most important peaks are recorded at the frequencies presented in Table IV. From the comparison of both cases results significant amplitude harmonics, common to the two types of windings, and namely the harmonics presented in Table V. III. CONCLUSIONS Modelling of the magnetic noise was performed at two speeds 1500 rpm and 3000 rpm using the finite element software FLUX 2D. It was studied the influence of magnetic pressure on each type of winding and it has been shown the significant harmonics which appear in this case presented in Table I,II,III and IV. Following this study results that for the analyzed motors can be used the simple layer winding without exceeding the noise level requested by standards. This study was performed with the purpose to reduce time and costs required to manufacture three-phase induction motor of type M3AA 71 B2 /4 through the practical implementation practice of the obtained solutions results, and as well to observe the transitory phenomenon that occurs during operation working. ACKNOWLEDGMENT We are grateful to Electroprecizia Sacele enterprise all the support given to accomplish the measurements for the present article. REFERENCES [1] Ionescu R. M., Contributions to modeling and analysis of the noise of variable speed induction machines, PhD Thesis, \u201cTransilvania\u201d University of Brasov , Romania, 2011 [2] Scutaru G., Peter I. The Noise of the electrical induction motors with squirrel-cage rotor (in Romanian language), Publishing house LUX LIBRIS, Bra\u015fov, 2004. [3] Ionescu R. M, Scutaru G. , Peter I, Analysis the magnetic noise of the induction motor with speed control, University Publishing House 'Transylvania' ,Brasov, 2011 [4] J. Gieras, C. Wang and J. Cho Lai, Noise of polyphase electric motors, Taylor&Francis, 2006. [5] J. Le Besnerais, V. Lanfranchi, G. Friedrich, M. Hecquet and P. Brochet, \u201cPrediction of audible magnetic noise radiated by adjustable speed drive induction machines\u201d, Proceedings of Electric Machines and Drives Conference, IEMDC 2009, Miami, Florida, USA, 2009, pp. 902-908. [6] Antti Laiho, \u201cElectromechanical modelling and active control of flexural rotor vibration in cage rotor electrical machines \u201d PhD Thesis,Helsinki University of Technology,Helsinki, 2009"
        ]
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.24-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.24-1.png",
        "caption": "Figure 3.24 Aircraft\u2019s plane of symmetry (the longitudinal plane) and the longitudinal motion, (\u03c3 = P = R = V = 0)",
        "texts": [
            "13), and (3.126)\u2013(3.128) is identified to be \u03c3 = P = R = V = 0, for which we have \u0307 = Q, \u0307 = 0, X \u2212 mg sin = m(U\u0307 + QW), Y = 0, (3.145) Z + mg cos = m(W\u0307 \u2212 QU), L = 0, M = JyyQ\u0307, N = 0, that is, the motion is confined to the plane of symmetry and can be entirely represented by the variables (t), U(t), W(t), Q(t), with = const. Such a motion has a special place in aircraft dynamics and is referred to as longitudinal motion. Consequently, the plane of symmetry,oxz, is also called the longitudinal plane (Figure 3.24). Navigational subsystem (b) for cruising flight of an aircraft essentially involves longitudinal motion wherein a constant airspeed and altitude must be maintained despite disturbances due to atmospheric gusts. The small rolling (banking) motion required for subsystem (a) can be neglected, leading to a flight confined to the plane of symmetry. Other examples of longitudinal dynamics are climb and descent, as well as take-off and landing. Longitudinal motion is normally represented by a small displacement from an equilibrium (unaccelerated) flight condition in the longitudinal plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003729_amr.487.401-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003729_amr.487.401-Figure2-1.png",
        "caption": "Fig. 2 Geometrical model of seal structure about double-Y",
        "texts": [
            " Finite element model 1.1 Geometric model Working conditions at the piston of rapping device: design pressure for 350 MPa, design temperature for 225 \u2103. The diameter of piston is 80mm. According to HG4-335-66 standard, choosing the size of Y-ring for 65 \u00d785\u00d7 10(Inner diameter \u00d7External diameter\u00d7 Height). According to the U.S. national standards, choosing the size of quad-ring for 65 \u00d77(Inner diameter\u00d7 Width). Geometric model and material parameters [5]of the two seal structure are shown in Figure 1, Figure 2 and Table 1. Fig.1 Geometrical model of seal structure about quad-ring 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: 128.119.168.112, Univ of Massachusetts Library, Amherst, USA-12/07/15,16:46:51) Table 1 Material properties Component Materials Species Modulus of elasticity/MPa Poisson\u2019s ratio \u00b5 Piston 3Cr13Mo 2\u00d710 5 0.3 Casing 3Cr13Mo 2\u00d710 5 0.3 Y-ring NBR 10 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003384_amm.344.3-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003384_amm.344.3-Figure2-1.png",
        "caption": "Fig. 2 Schematic plan of joint wear calculation of ACSPB under axial load",
        "texts": [
            " If the limiting wear intensity H is given, the wear-life of the ACSPB can be calculated as 1 2 )( m r r rL H Hf nKP   (12) Wear-life model for ACSPB under axial load Just as the process of getting the wear-life model for ACSPB under radial load, there were also two simplifications: Firstly, the load acting on the ACSPB was mainly the thrust load (pure thrust load or the composite load of radial load and thrust load while the thrust load was the chief load). The thrust load acted on ACSPB along the axial lead. The second simplification was the same as the one for ACSPB under radial load. The direction of joint wear of ACSPB was along the x-x direction according to the working condition above, as shown in Fig.2 (the section in figure was defined by point G and the axial lead). According to Fig.2, the relative sliding velocity at point G can be achieved as 2v nRsin  (13) From Eq.(2) and (13), the wear velocities at point G in outer ring and inner ring can be defined as 1 1 12m mk p v nRsin k p    , 2 2 22m mk p v nRsin k p    (14) According to the contact qualification of the JWCM [6] , the joint wear rate \u03b31-2 can be defined as 1 2 1 2 1 2( 2) / /mcos nR k k p sin cos          \uff08 \uff09 (15) Then the contact pressure p can be achieved as 1 1 2 1 2/(2 ) mp cos nR k k sin    [ \uff08 \uff09 ] (16) Considering the relationship between the contact pressure and the load, the load Pa is 2 2 1 1 2 2 1 1 2 1 1 2 1 1 2 1 1 1 1 11 2 1 2 1 2 1 1 1 2 2 2 2 2 /(2 ) ( ) ( ) a a m m m m m m m m P p cos dS p Rsin Rd cos R psin cos d P R R sin cos cos nR k k sin d sin cos d n k k                                                    \uff08 \uff09 [ \uff08 \uff09 ] \uff08 \uff09 (17) 6 Advanced Research on Applied Mechanics, Mechatronics and Intelligent System The joint wear rate \u03b31-2 can be achieved as 1 2 1 1 2 1 2 2m m m a aP n k k R I      ( ) / ( ( ) ) (18) Where 2 1 1 1 1 1( )m m m aI sin cos d       ( ) ( ) (19) \u03b81 and \u03b82 are the configuration parameters of ACSPB, 1 arccos(( ) )T s C R    "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003012_12.908771-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003012_12.908771-Figure1-1.png",
        "caption": "Figure 1. Technical drawing of the fiber sensor consisting of two silver halide fibers positioned on a custom-made titan alloy holder.",
        "texts": [
            " The fibers were positioned on a custom-made titan alloy holder (6AL4V, THG Titan-Halbzeug GmbH, Mo\u0308nchengladbach, Germany) with a cross-sectional width of 1.5 mm and a length of 4 cm. Both fibers were fixed with a biocompatible adhesive (Loctite 4061, Henkel AG & Co. KGaA, Du\u0308sseldorf, Germany). The gap between the fibers served as a measurement cavity with a path length of approximately 50 \u00b5m. The total length of the sensor amounted to 16 cm. A technical drawing of the fiber sensor can be found in Fig. 1. Mid-IR radiation originated from two Fabry-Perot QCLs (Alpes Lasers SA, Neucha\u0302tel, Switzerland) which operated at wavenumbers 1035 cm\u22121 (corresponding to a wavelength of 9.66 \u00b5m, QCL1) and 872 cm\u22121 (11.47 \u00b5m, QCL2), respectively. QCL1 ran in pulsed mode with a pulse width of 50 ns and at a repetition period of 3.3 \u00b5s (duty cycle 1.5%) resulting in an average output power of 1.8 mW. QCL2 was driven with a pulse width of 50 ns and at a repetition period of 1 \u00b5s (duty cycle 5%) with an average output power of 6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002137_978-1-4614-1150-5_12-Figure12.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002137_978-1-4614-1150-5_12-Figure12.1-1.png",
        "caption": "Fig. 12.1 An object subjected to externally applied forces",
        "texts": [
            " The second type of motion involves local changes of shape within a body, called deformations, which are the primary concern of the field of deformable body mechanics. If a body is subjected to externally applied forces and moments but remains in static equilibrium, then it is most likely that there is some local shape change within the body. The extent of the shape change may depend upon the magnitude, direction, and duration of the applied forces, material properties of the body, and environmental conditions such as heat and humidity. Consider the arbitrarily shaped object illustrated in Fig. 12.1, which is subjected to a number of externally applied forces. Assume that the resultant of these forces and the net moment acting on the object are equal to zero. That is, the object is in static equilibrium. Also assume that the object is fictitiously separated into two parts by passing an arbitrary plane ABCD through the object. If the object as a whole is in equilibrium, then its individual parts must be in equilibrium as well. N. O\u0308zkaya et al., Fundamentals of Biomechanics: Equilibrium, Motion, and Deformation, DOI 10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003269_s1068798x11080120-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003269_s1068798x11080120-Figure1-1.png",
        "caption": "Fig. 1. Loading of teeth in transmission. Fig. 2. Model (a) of the contact between teeth 1 and 2, without (b) and with (c) skewing: K, point of tooth con",
        "texts": [
            "5Fn/ \u03c0aloblo( ),= \u03c3lo.cl.e 2 0.5\u2013 \u03c3x \u03c3y\u2013( ) 2 \u03c3y \u03c3z\u2013( ) 2 \u03c3z \u03c3x\u2013( ) 2 6 \u03c4xy 2 \u03c4yz 2 \u03c4zx 2+ +( )+ + += \u03c3lo.cl.e 1 2\u03bc\u2013( )Zlo.s\u03c3lo.cl,= \u03b2 Zlo.s 1 \u03b2\u2013 \u03b2 2 +( ) 0.5 / 1 \u03b2+( ).= DOI: 10.3103/S1068798X11080120 726 KOROTKIN, GAZZAEV contact area moves toward one end of the crown; the contact stress increases here. We simulate the stress\u2013strain state of the teeth on the basis of the finite element method using ANSYS software, by solving the three dimensional contact problem. The loading of the teeth is shown in Fig. 1. Teeth 1 and 2 are rigidly fixed at the bases, with corresponding determination of the boundary conditions and replacement of the missing parts of the gear rims; the attachment parameters are selected on the basis of the results in [6]. Application of a rotating torque allows the teeth to experience not only contact but flexural\u2013 shear deformation relative to the rigid attachment. The rotating torque Pl relative to the O2 axis is applied to tooth 2 through its attachment; the necessary normal force Fn corresponds to the relation FnR2cos\u03b1 = Pl, where R1 and R2 are the initial radii of gears with teeth 1 and 2, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002825_isie.2013.6563881-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002825_isie.2013.6563881-Figure7-1.png",
        "caption": "Fig. 7 Holes drilled on the rotor to avoid the current flow through the bar.",
        "texts": [
            " By the simple connection of a 3 phase AC mains voltage (400 VL-L), a DC-Bus capacitor (1100\u03bcF) and the IGBTs control signals circuitry, a complete power stage is achieved. The heat sink was carefully selected to provide sufficient creepage and clearance distances; a fan was also attached to cool down the component. The control and fault signals are optically isolated with the power stage to avoid undesired current peaks in case of a failure. The parameters of the motor are listed in the Appendix at the final part of this document. As shown in Fig. 7, two holes were drilled to completely avoid the current flow through one of the 28 bars of the squirrel cage rotor. To prove the effectiveness of the identification method, several experiments were carried out and are presented on this section. An alternating voltage was applied to the stator windings at a frequency of 400 Hz and 300 Vpk. Instead of measuring the admittance for different rotor positions as explained above, the machine was not fed by a single phase supply as in Fig. 3 but with a three phase inverter so the direction of the applied voltage phase phasor could be rotated and fixed to a determined axis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.42-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.42-1.png",
        "caption": "Fig. 2.42 4PRPaPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology P\\R||Pa\\Pa",
        "texts": [
            " 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure7-1.png",
        "caption": "Fig. 7 Details of the structure",
        "texts": [
            " Stricture Scheme A new kind of quick connector was designed based on the structure of a push-push type latch mechanism and electronic device thereof [9]. This machine was designed for a connection of a function bed and a devise which hold a tubular shape medical device, a catheter [10] for example. Fig.6 shows a full view of the connector. It includes a holding device which is the male part of the connection, and the locking part links with the function bed, which is the female part of the connection. Details of the structure are shown in Fig.7. The process of this connection will be descript as following: firstly, push the holding device down to the base plate, then the side wall of the holding device will press the locking pin, top of the locking pin moves on the surface of the holding device. And the locking pin presses the locking spring, this situation goes on until the locking pin goes into the groove in the holding device (doesn\u2019t shown in the picture), and now the locking spring is still pressed, it presses the locking pin, and the locking pin is pushed to the wall of the function bed, this makes friction force to lock the holding device and the function bed, connection accomplished. In Fig.7 we can see the separation button is in the shape of wedge, the wedge top goes to a groove in the locking pin as the size of the wedge. If the operator presses the separation button, the locking springs will be pressed, the separation button will be pushed into the groove of the locking pin, the surface of the wedge will push the groove interface, and then the locking pin will be moved out of the groove in the holding device, so there will be no pressure on the other surface of the holding device, then the friction force disappears"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure8-1.png",
        "caption": "Fig. 8: Bottom Frame Reaction Analysis",
        "texts": [],
        "surrounding_texts": [
            "The authors like to acknowledge the PACE program and GM for sponsoring this project. The authors also like to thank the faculty member and students of Hongik University (South Korea) for helpful discussions in this project."
        ]
    },
    {
        "image_filename": "designv11_100_0002970_amm.86.619-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002970_amm.86.619-Figure3-1.png",
        "caption": "Fig. 3 Finite element gridding",
        "texts": [
            " Only a right selection can ensure a most proximity to simulated actual working conditions. A three-dimensional model is built (as shown in Fig. 1). Force analysis has two steps. Step 1: Cohesive joints are selected and their rigidity matrix is worked with ANSYS. Step 2: The matrix is introduced into MASTA to work out the results of forces on each cohesive joint from loads, as shown in Fig. 2 below. Housing is gridded into 8-joint-and-6- surface units with FEA. It follows 2 steps. Step 1: general gridding and calculation. Step 2: detailed gridding (as shown in Fig. 3) based on force situation. The stronger forces are, the closer grids are. The final results (as shown in Fig. 4-a & Fig. 4-b) have been calculated: unit total \u2013 206,149; joint total \u2013 277,989. As calculated, the maximum equivalent stress on housing (as shown in Fig. 4-a) is 104MPa and the maxim distortion by comprehensive displacement (as shown in Fig. 4-b) is 0.11mm, which indicate that distortion appears in good distribution and its rigidity meets the design standard. As presented above, this design of wind turbine gearbox housing, which takes pertinent measures in readjusting its housing structure so as to overcome the weak points detected with FEA, is an optimal one to achieve low weight, high quality and high safety"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003320_ijaac.2011.043611-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003320_ijaac.2011.043611-Figure2-1.png",
        "caption": "Figure 2 Reference frames and parameters of RDIP",
        "texts": [
            " In Sections 5 and 6, we discuss the concepts in details and illustrate them with simulations. Section 7 concludes this research. The RDIP system consists of a controller, an arm, two pendulums, an actuators (DC motor) and three increment rotary encoders. The controller makes the two pendulums stand at upright position on the rotary arm by moving the arm on the base. The motor provides power to rotate the arm. The encoders detect the pendulums and arm angular position as shown in Figure 1. Reference frames and some parameters of the RDIP are defined as shown in Figure 2. The variables of 1 , 1 and 1 are angular position, velocity and acceleration of link 1, respectively, which is the arm; 2 , 2 and 2 are angular position, velocity and acceleration of link 2, respectively. 3 , 3 and 3 are angular position, velocity and acceleration of link 3, respectively. 1 2,J J and 3J represent moments of inertia of the three respective links about their centre of mass. 1 2,l l and 3l are the distances from the centre of rotation of the links to the centre of mass of the respective links",
            " Derivation of mathematical equation describing dynamics of the RDIP system is based on Euler\u2013Lagrange equation of motion given by: d , 1, , d i i i i L L W Q i m t q q q (1) where ( )q t is the angular position vector, ( )q t is the angular velocity vector, iQ is the external force or load vector, L is the Lagrangian and W is the loss energy. In the Euler\u2013 Lagrange equation, the Lagrangian L is defined as: total total( , )L q q T V (2) where totalT is total kinetic energy of the RDIP system and totalV is total potential energy of the RDIP system. From Figure 2, the total kinetic energy of the system is the sum of kinetic energy of the rotary arm and the two pendulums given by: total arm rod1 rod2T T T T (3) Kinetic energy of link 1, the rotary arm is: 2 2 2 arm 1 1 1 1 1 1 1 2 2 T m l J (4) When 1 0l , since the centre of mass of the arm is balanced at the original point, thus: 2 arm 1 1 1 2 T J (5) Similarly, the kinetic energy of links 2 and 3 are determined as: 2 2 2 2 rod1 2 1 1 2 2 2 2 2 1 1 2 2 2 1 1 cos 2 2 T m L m l J m L l (6) 2 2 2 2 2 2 rod2 3 1 1 3 2 2 3 3 3 4 3 1 2 1 2 2 3 1 3 1 4 4 3 2 3 2 4 2 4 1 1 1 2 2 2 T m L m L m l J m L L cos m L l cos m L l cos (7) The total potential energy for the system is the sum of potential energy of the rotary arm and the two pendulums is given by: total arm rod1 rod2V V V V (8) total 2 2 2 3 2 2 3 40 cos cos cosV m gl m g L l (9) The total loss energy of the system is the sum of loss energy of the rotary arm and the two pendulums given by: 2 2 2 1 1 2 2 3 4 1 1 1 2 2 2 W C C C (10) From Equation (2), LaGrange is determined as: arm rod1 rod2 rod1 rod2L T T T V V (11) Substitution of Equations (5), (6), (7) and (9) into Equation (11) yield: 2 2 2 2 2 1 1 2 1 1 2 2 2 2 2 1 1 2 2 2 2 2 2 2 2 2 3 1 1 3 2 2 3 3 3 4 3 1 2 1 2 2 3 1 3 1 4 4 3 2 3 2 4 2 4 2 2 2 3 2 2 3 4 1 1 1 cos 2 2 2 1 1 1 cos 2 2 2 cos cos cos ( cos cos ) L J m L m l J m L l m L m L m l J m L L m L l m L l m gl m g L l (12) Euler\u2013Lagrange equation of the motion of each variable, thus becomes: 11 1 d d e L L W t (13) 22 2 d 0 d L L W t (14) 44 4 d 0 d L L W t (15) Substitution of Equation (12) into Equations (13)\u2013(15) and solving of Euler\u2013Lagrange equation yield: 2 2 1 2 1 3 1 1 2 1 2 3 1 2 2 2 3 1 3 4 4 2 2 2 1 2 3 1 2 2 2 3 1 3 4 4 1 1 cos cos sin sin e J m L m L m L l m L L m L l m L l m L L m L l C (16) 2 2 2 1 2 3 1 2 2 1 2 2 2 3 2 2 3 2 3 2 4 4 2 3 2 3 4 2 4 2 2 3 2 2 2 2 cos cos sin sin 0 m L l m L L J m l m L m L l m L l m l g m L g C (17) 2 3 1 3 4 1 3 2 3 2 4 2 3 3 3 4 2 3 2 3 2 4 2 3 3 4 3 4 cos cos sin sin 0 m L l m L l J m l m L l m l g C (18) To simplify the equations, some parameters are defined as shown in Table 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002656_wcse.2013.52-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002656_wcse.2013.52-Figure3-1.png",
        "caption": "Fig. 3. blade element frame of axes.",
        "texts": [
            " (2) Q=CQ A( R)2 . (3) H=CH A( R)2 . (4) CT, CD, CQ and CH are aerodynamic coefficients, means angular speed of rotor, R is rotating radius of rotor. In order to calculate conveniently, two important velocity ratios, inflow ratio and advance ratio , are given by, . (5) . (6) 978-1-4799-2883-5/13 $31.00 \u00a9 2013 IEEE DOI 10.1109/WCSE.2013.52 283 The component of V0 along xsoszs plate has different effects on different positions of blades. For analyzing easily, blade element frame of axes op-xpypzp is shown as Fig.3. Define as azimuth of blade, means that blade is located in the negative direction of xs axis. Select blade element dr, whose average chord length is c, and the distance from which to os is r. The blade element is shown as Fig.4. Where , represent attack and inflow angles of blade respectively, means setting angle, the relationship among , and is, \u2248 . (7) Vp and Vt are vertical and circumferential components of V0, whose values are given by, . (8) When neglecting the effect of radial airflow, blade element develops derivative lifting force dL and derivative drag dD, whose directions are vertical and parallel to the direction of airflow, and whose values are given by, Where is air density, means lifting coefficient, , a is constant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002110_978-3-319-01845-4_6-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002110_978-3-319-01845-4_6-Figure1-1.png",
        "caption": "Fig. 1 Kinematic chain",
        "texts": [
            "\u2019\u2019 \u2018\u2018System of bodies designed to convert motions of, and forces on, one or several bodies into constrained motions of, and forces on, other bodies (Ionescu 2003) (see footnote 1).\u2019\u2019 The structural parameter of the kinematic chain and mechanism are defined in (Ionescu 2003) (see footnote 1) as: \u2018\u2018The degree of freedom of a kinematic chain or the degree of freedom (mobility) of a mechanism represents the number of independent coordinates needed to define the configuration of a kinematic chain or mechanism.\u2019\u2019 The degree of freedom of a kinematic chain (Fig. 1) in respect with a reference system may be calculated according to (Dobrovolski 1951) as: L \u00bc 6n X5 i\u00bc1 i ci ; \u00f01\u00de where ci is the number of kinematic pairs of class i, i = 1,2,\u20265; n\u2014the number of links of the kinematic chain. The relationship describing the degree of freedom (mobility) of mechanism is: M \u00bc \u00f06 f \u00de \u00f0n 1\u00de X5 i\u00bcf\u00fe1 \u00f0i f \u00de ci; \u00f02\u00de where f represents the number of common restricted motions of the mechanism\u2019s elements usually named \u2018\u2018the family\u2019\u2019 (Manolescu 1968; Antonescu 1973; Kov\u00e1cs et al"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003778_978-3-642-17234-2_4-Figure4.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003778_978-3-642-17234-2_4-Figure4.13-1.png",
        "caption": "Fig. 4.13 Motion of the symmetrical top in the gravitational field.",
        "texts": [
            "; w \u00bc X0t: \u00f04:6:11\u00de Observation: The geometric locus of the instantaneous axes of rotation with respect to S0 is a cone with its top at O, called polhodic cone; the geometric locus of the instantaneous axes of rotation relative to S is also a cone with the top at the same point, named herpolhodic cone. It can be shown that the two cones are tangent, and the polhodic cone is rolling without slipping over the herpolhodic cone (Fig. 4.12). Suppose that the fixed point of the top is on its axis of symmetry. We choose the origins O and O0 in the fixed point, and the axis Oz0 directed along the symmetry axis of the top. As one can see, in this case the centre of mass and the fixed point do not coincide anymore (Fig. 4.13). Then, if the principal moments of inertia ~I1;~I2;~I3 relative to a frame fixed with respect to the rigid and having its origin in G are known, from (4.5.9) we obtain: I01 \u00bc I02 \u00bc ~I1 \u00feMl2; I03 \u00bc ~I3; \u00f04:6:12\u00de where M is the mass of the top and l is the distance between the fixed point O and the centre of mass G. or, in view of (4.5.2), T \u00bc 1 2 I01\u00f0 _u2 sin2 h\u00fe _h2\u00de \u00fe 1 2 I03\u00f0 _u cos h\u00fe _w\u00de2: \u00f04:6:13\u00de The potential energy of the body situated in the terrestrial gravitational field is V \u00bc Mgl cos h: The Lagrangian is then: L \u00bc 1 2 I01\u00f0 _u2 sin2 h\u00fe _h2\u00de \u00fe 1 2 I03\u00f0 _u cos h\u00fe _w\u00de2 Mgl cos h: \u00f04:6:14\u00de The two cyclic coordinates u;w lead to the following first integrals: pu \u00bc oL o _u \u00bc \u00f0I01 sin2 h\u00fe I03 cos2 h\u00de _u\u00fe I03 _w cos h \u00bc C1; \u00f04:6:15\u00de pw \u00bc oL o _w \u00bc I03\u00f0 _u cos h\u00fe _w\u00de \u00bc I03x 0 3 \u00bc C2: \u00f04:6:16\u00de Since the physical system is conservative, there also exists the energy first integral: 1 2 I01\u00f0 _u2 sin2 h\u00fe _h2\u00de \u00fe 1 2 I03\u00f0 _u cos h\u00fe _w\u00de2 \u00feMgl cos h \u00bc E: \u00f04:6:17\u00de The constants C1;C2;E are determined from the initial conditions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003549_ptc.2013.6652096-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003549_ptc.2013.6652096-Figure1-1.png",
        "caption": "Figure 1. Squirrel cage induction generator (SCIG) model.",
        "texts": [],
        "surrounding_texts": [
            "Index Terms-- Electric Power Systems, Electromechanical Transient Analysis, Sliding Mode Controllers, Wind Turbines.\nI. INTRODUCTION The constant increase in electricity demand has led to a diversification of the energy supply, this, associated with the increasing interest in clean energy sources, makes most of the countries seek for new technologies that could meet these needs. In this way, there has been a large increase in the number of wind turbines connected to power system. In Brazil, for example, it is expected that the installed capacity will be increased from 831MW to 11.532MW by 2020 [1].\nConsidering the technological development and consequent increase in rated capacity, the impact of integrating wind turbines with the network has been considered increasingly important for both steady state and dynamic conditions.\nThis increase together with the inclusion of wind turbines have a direct influence on the quality of service and may interfere with the stability of the entire electrical system, since the amount of daily energy available can vary greatly from one season to another. Also, its use can be limited to places of high and relatively constant winds.\nDoubly-fed induction generator (DFIG) based wind turbines have been one of the most emergent technologies for this kind of energy generation, since this kind of asynchronous machine is a cost effective, efficient and reliable solution and, when compared with fixed speed induction squirrel-cage induction generators and synchronous generators, it has some advantages due to its variable speed operation and four quadrant active and reactive power capabilities [2].\nThe characteristics of operation and control of wind turbines and their impact on the stability of the electrical system could be found in references such as [3-6].\nThus, computational tools that allow the simulation of the behavior of the electrical system in steady state and dynamic condition, considering the use of wind turbines, are used.\nTo evaluate the behavior of wind turbines, a power control scheme for a grid-connected DFIG-based wind farm, using stator flux decoupling and sliding mode controllers, was developed.\nThereafter, this paper proposes a dynamic analysis of the impacts, especially electromechanical transients caused by DFIG-based wind turbines connected to the transmission electrical systems.\nII. WIND TURBINE GENERATORS The energy conversion of most modern wind turbines can be divided into fixed speed or variable speed. The fixed-speed induction generator is represented by squirrel cage induction generator (SCIG). The DFIG and variable speed direct drive synchronous generator (DDSG) represent the main types of generators used in a variable speed unit, which is connected to the grid using the power electronics converter technology. Figs. 1-3 illustrate these configurations [7].",
            "III. DOUBLY FED INDUCTION GENERATOR MODEL The doubly-fed induction generator model is given by [8]:\ndq dq\ndqdq j dt\nd iRv 11 1 111 (1)\ndqmec dq\ndqdq PPj dt\nd iRv 21 2 222 )( (2)\nthe relationship between fluxes and currents are:\ndqMdqdq iLiL 2111 (3)\ndqdqMdq iLiL 2212 (4)\nThe machine dynamics is given by:\nMdqdq mec TiPP dt d J )Im( * 112 3 (5)\nThe generator active and reactive power are:\n)( 2 3 1111 qqdd iiP (6)\n)( 2 3 1111 qddq iiQ (7)\nThe subscripts 1 and 2 represent the stator and rotor parameters respectively, 1 is the synchronous speed, mec is the machine speed, R1 and R2 are the per phase resistances of the stator and rotor windings, L1, L2 and LM are the self and mutual inductances of the stator and rotor windings, v is the\nvoltage vector, i is the current vector, is the flux vector , PP is the machine number of pair of poles, J is the load and rotor inertia moment and TM is the mechanical torque.\nThe DFIG power control aims independent stator active P and reactive Q power control by means of the rotor current regulation. For this purpose, P and Q are represented as functions of each individual rotor current. Using the stator flux oriented control, that decouples dq axis (3) becomes:\nd M d i L L L i 2\n11\n1 1 (8)\nq M q i L L i 2 1 1 (9)\nWhere 1d = 1 = dq1 . The active (6) and reactive (7) power\ncan be calculated using (8) and (9):\nq M i L L P 2 1 12 3\n(10)\nd M i L L L Q 2\n11\n1 12\n3 (11)\nWhere v1 = v1q = | dqv1 |. Thus, the rotor currents will be reflected in the stator currents and on the stator active and reactive power. Consequently this principle can be used on the stator active and reactive power control of the DFIG.\nTo this work, an alternative power control scheme for DFIG using Sliding Mode Controllers (SMC) with stator-fluxoriented vector control is proposed.\nIV. SMC APPLIED TO THE DFIG POWER CONTROL The essential idea of traditional VSC control algorithms is to enforce the system mode to slide along a predefined sliding surface of the system state space [10].\nOnce the state of the system reaches the sliding surface, the structure of the controller is adaptively changed to slide the state of the system along the sliding surface. Hence, the system response depends only on the predefined sliding surface and remains insensitive to variations of system parameters and external disturbances. However, such insensitivity property is not guaranteed before the occurrence of the sliding mode, resulting in loss of robustness while the control system is reaching the phase.\nFurthermore, in order to reduce the chattering effect, the sign function of the VSC is often replaced by a saturation function in practical implementations.",
            "In [11] a sliding mode approach for direct torque control of sensorless induction motor drives is proposed. Reference [12], proposes some controllers for field orientation control (FOC) based on the dynamic modeling of the DFIG.\nThe strategy herein proposed uses the sliding mode controller and the stator-flux- oriented control to regulate the rotor currents and the active and reactive power based on (10) and (11).\nThe error between the current references and the measured values to obtain the sliding surface can be defined as:\nddrefdi iie 222 - (12)\nqqrefqi iie 222 - (13)\nWhere i2d and i2q are the rotor currents calculated on the dq referential frame. i2dref and i2qref are the rotor current references computed by PI controllers that process the errors between the reference and measured power in accordance with (10) and (11). The sliding surface S can be defined as:\n)(\n)(\n222\n222\n2\n1\nqiqiqi\ndididi\ne dt dce\ne dt dce\nS S S (14)\nWhere ci2d and ci2q are constants defined taking into account the desirable dynamic response for the system. According to (10) and (11), the d current rotor component is responsible for the reactive power control whereas, the q current rotor component is responsible for the active power control. This way, the control objective is to make the system state go to the equilibrium point defined on the origin of the sliding surface (S = 0), where the errors and their derivatives are zero, ensuring that the states reach their references. Based on [11], for active and reactive power control by regulating rotor currents, the rotor voltage references are given by:\n)().(K 1 2 Pi2d2 seval s\nK v dIi\ndref (15)\n)().(K 2 2\nPi2q2 seval s\nK v qIi\nqref (16)\nWhere KPi2d and KPi2q are the proportional gains and KIi2d and KIi2q are the integral gains of the PI controller; v2dref and v2qref are the rotor voltage references on the dq reference frame; and, eval(s1) and eval(s2) are evaluation functions that determine the switching behavior of the controller once the responses reach the sliding surface. The eval function can be simple as the signal function. However in this case, it was used a saturated linear function, as given by:\n.max.min, max;.min,.\nmax;.max, )(\nn\nnn\nn\nn sKif sKifsK\nsKif seval (17)\nWhere n can be 1 or 2, depending on the sliding surface to be used in (14).\nThis way, the control algorithm generates the rotor voltage references that allow the active and reactive power to converge to their respective reference values. The desired rotor voltage in r (the rotor reference frame) that generates switching signals for the rotor side converter using, for example, space vector modulation, is given by:\nrsevv dqr 22 (18) The block diagram of the proposed control scheme is\nshown in Fig. 4.\nFigure 4. Control system used.\nTo accomplish with new grid code requirements some methodologies have been proposed in the literature. As the stator of the DFIG is connected direct by to the grid, some undesirable high currents may be induced in the rotor windings and the protection system may block the RSC.\nAnother undesirable transient is the voltage at the DC-link, which can reach high levels depending on two main characteristics: low residual terminal voltage during fault and the slowness of the RSC disconnection from the rotor winding after fault detection. The strategy is based on the current compensation injection on the DFIG control during the fault [13].\nThis strategy, presented in [13], aims to reduce the currents in the stator/rotor windings when a fault occurs, because they can damage the DFIG. The idea is to feedback the measured stator currents as current references for the controller of RSC when a voltage sag occurs. This is done based on the relationship shown in (19) and (20) [13].\nd m qd i L L v sLRsL i 2 1 12 111 2 1 1 1 1 )/(2L 1 (19)\nq m qq i L L v sLRsL LRs i 2 1 12 111 2 1 11 1 1 )/(2 )/( L 1 (20)"
        ]
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure6-1.png",
        "caption": "Figure 6: Components of TMP holder",
        "texts": [
            "org/terms TMP manipulation stage is used for holding, moving and adjusting TMP. The stage is comprised of XYZ liner stages and TMP holder (Figure 5). The 3-axis liner stage with a repeatable positional accuracy of 0.5 \u03bcm can control the TMP holder motion in the XYZ direction. TMP holder is consisted by vacuum chuck, magnetic basement, 3 axis inclined stage and multi-axis force/torque sensor. The method of fixing the vacuum chuck and magnetic basement is based on magnetism and 3 axis \u201cV\u201d fix groove (Figure 6). This method can meet precise repeatable position requirement and fix quickly. There are TMP supporting ring and TMP locating pins on the work plane of the vacuum chuck. TMP supporting ring can protect TMP in assembly process. TMP locating pins can guarantee the axis of TMP and TMP supporting ring to be concurrent. The 3-axis inclined stage is located between 6-Axis Force/Torque sensor and Magnetic fix basement. It can adjust the position error of TMP. In the assembly process, the contact force of TMP and Hohlraum can be detected by 6-axis force / Proc"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure12-1.png",
        "caption": "Figure 12: Finished half Hohlraum component",
        "texts": [],
        "surrounding_texts": [
            "The first experiment which is carried out on Hohlraum inserted into TMP by the prototype system is successful, and the finished Half Hohlraum Component is shown in Figure 13. The friction force of Hohlraum inserting is 3 N which is greater than the theoretical value of 0.5 N in the experiment. The parallelism tolerance of manual stage, mismachining tolerance and attitude tolerance of the subassemblies may be the reasons of making the experimental friction force greater than the theoretical one. Nevertheless, the torn surfaces of Hohlraum can not be observed by microscope. The experiment result shows that the non-coaxial measurement of Hohlraum and TMP is 8 \u03bcm, which is larger than NIF target requirement of 2 \u03bcm. The first experiment results verify the feasibility of Hohlraum assembly."
        ]
    },
    {
        "image_filename": "designv11_100_0003081_gt2012-69967-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003081_gt2012-69967-Figure3-1.png",
        "caption": "Figure 3. Geometric definition for the square/hexagon method from Nestorides [1].",
        "texts": [
            " 4 32 o d J   (1) o o J G K l  (2) As described previously, Nestorides [1] provides four methods for determining the torsional stiffness of a shaft with radial webs. The first method assumes a square or hexagon cross section that encompasses the base shaft diameter and overlaps into the cross-section of the webs. The torsional stiffness of the webbed shaft is approximated by calculating the torsional stiffness of the shaft with the assumed polygon crosssection. A square cross-section is applied for a four-webbed geometry and a hexagon is assumed for a six-webbed configuration, as shown below in Fig. 3. For shafts with polygon cross-sections, equivalent diameters can be assumed, allowing for a torsional stiffness calculation to follow the same method as a circular shaft provided in Eqs. (1) and (2). The torsional equivalent diameters for square and hexagon crosssections are provided in Eqs. (3) and (4), respectively. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 4 Copyright \u00a9 2012 by ASME   1/4 , 1.431eq squared d (3)   1/4 , 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure1-1.png",
        "caption": "Figure 1 FDM drive system",
        "texts": [
            " FDM is an extrusion-based process that feeds material in solid wire form and then melts it into a shape and forms a solid. FDM is a non-laser filament extrusion process that utilises engineering thermoplastics, which are heated from filament form and extruded in very fine layers to build each model from the bottom up. The models can be made from acrylonitrile butadiene styrene, polycarbonate, polyphenylsulphone (PPSF), Ultem and various versions of these materials. Furthermore, in many cases, the models are tough enough to perform functional tests. Figure 1 is a depiction of the head and process in which the material is pulled off the spool, heated just above the melting temperature and deposited at the desired location. The key steps are as follows: 1 starting with the filament being fed into the drive wheels 2 the drive wheels force the filament into the liquefier 3 the heater block melts the filament 4 the solid filament is used as a linear piston 5 the melted filament is forced out through the tip. Source: Liou (2007). The material used is fed into the head in solid wire form and then liquefied in the head and deposited through a nozzle in liquid form"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.1-1.png",
        "caption": "Figure 10.1 An induction motor: (a) rotor and stator assembly; (b) rotor squirrel cage; and (c) cross-sectional view of an ideal induction motor with six conductors on the stator",
        "texts": [
            " However, when compared to PM motors, induction motors have lower efficiency and less torque density. Typical induction motors used for traction applications are squirrel cage induction motors. An inverter is used to control the motor so that the desired torque can be delivered for a given driving condition at a certain speed. Advanced control methodologies, such as vector control, direct torque control, and field-oriented control, are popular in induction motor control for traction applications. The basic structure of an induction machine is shown in Figure 10.1. The two main parts of an induction motor are the stator (which houses the winding) and the rotor (which houses the squirrel cage). Both stator and rotor are made out of laminated silicon steel with thickness of 0.35, 0.5, or 0.65 mm. The laminated steel sheets are first stamped with slots and are then stacked together to form the stator and rotor, respectively. Windings are put inside the stator slots while the rotor is cast in aluminum. There are some additional components to make up the whole machine: the housing that encloses and supports the whole machine, the shaft that transfers torque, the bearing, an optional position sensor, and a cooling mechanism (such as a fan or liquid cooling tubes). In Figure 10.1c, AX is phase a, BY is phase b, and CZ is phase c. The direction of the phase currents is for a particular moment \u03c9t = 60 electric degrees; \u201c+\u201d indicates positive and \u201c\u2013\u201d indicates negative. It can be seen that conductor AZB forms one group and XCY forms another group. Together they create a magnetic field at 30\u25e6 NW\u2013SE. The direction of the field will change as the current changes over time. The stator windings shown in Figure 10.1c are supplied with a three-phase AC sinusoidal current. Assume the amplitude of the currents is Im amperes, and the angular frequency of the current is \u03c9 radians per second; then the three phase currents can be expressed as ia = Im cos (\u03c9t) ib = Im cos (\u03c9t \u2212 120\u25e6) (10.1) ic = Im cos (\u03c9t \u2212 240\u25e6) Since the currents of each of the three phases are functions of time, the direction of current as shown in Figure 10.1c will change with time. If we mark the direction of the current at any given time, we can see the magnetic field generated by the stator current with its peak changing position over time. Mathematically, we can derive this magnetic field. Each of the three-phase currents will generate a magnetic field. Since the three windings are located 120\u25e6 from each other in space along the inside surface of the stator, the field generated by each phase can be written as follows, assuming the spatial magnetic field distribution in the air gap due to winding currents is sinusoidal by design: Ba = Kia (t) cos (\u03c9t) Bb = Kib (t) cos (\u03c9t \u2212 120\u25e6) (10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.18-1.png",
        "caption": "Figure 10.18 Temperature distribution in stator",
        "texts": [
            " In order to understand the impact of switching frequency on the additional losses, tests have been conducted using several different switching frequencies (3.5\u20139 kHz) for the PWM inverter. The results are shown in Figure 10.17. The total losses decrease proportionally with the increase of frequency, which is consistent with prior work [41]. By assigning all the losses into the 3D FEM model and incorporating the necessary boundary conditions based on the model of an actual motor, the temperature profile can be obtained. Figure 10.18 shows the temperature distribution in the stator with no load in sinusoidal supply. Tables 10.3 and 10.4 show the entire temperature of the motor, and the location of test points as shown in Figure 10.19 except point 1, which is in the center of the outer surface housing. It can be seen from Tables 10.3 and 10.4 that the temperature rise of the induction motor is significantly higher when driven by the PWM inverter than by sinusoidal power supply. In conclusion, for induction motor design it is important to consider the additional losses generated due to the harmonics in the PWM supply and possible Temperature [\u00b0C] excessive temperature rise inside the machine"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003067_amr.199-200.824-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003067_amr.199-200.824-Figure3-1.png",
        "caption": "Fig. 3 Symmetric spring characteristic model",
        "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. (ID: 128.6.218.72, Rutgers University Libraries, New Brunswick, USA-01/06/15,01:17:29) A without preloading condition is considered where the mass starts vibration from the midpoint of clearance. It is only possible when a reasonable stiffness spring exists between the driving side coupling part and the driven side coupling part. The system without preloading is denoted by symmetric model in this study. Preloaded model is discussed later in this literature. Fig. 3 shows the characteristic models of the spring k1, k2 and k3 where the vibration starts from O. Mathematical Modeling. The forces involve for the mass 1m of the symmetric two degree of freedom motion are: ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( )           \u2212>\u2212\u2212+\u2212+\u2212++\u2212+\u2212+\u2212 >\u2212\u2212+\u2212+\u2212+\u2212\u2212+\u2212+\u2212 \u2264\u2212\u2264\u2212\u2212+\u2212+\u2212+\u2212 = )(........)( ).....(....)( )(........................................)( ),( 01212212013013011011 01212212013013011011 01212212011011 xxdxxcxxkxxcdxxkxxcxxk dxxxxcxxkxxcdxxkxxcxxk dxxdxxcxxkxxcxxk xxf The forces involve for the mass 2m of the symmetric two degree of freedom motion are: ( ) ( )122122),( xxcxxkxxF \u2212+\u2212= for all conditions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003644_amr.605-607.1176-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003644_amr.605-607.1176-Figure2-1.png",
        "caption": "Fig. 2 The model of eccentric shaft after constraint",
        "texts": [
            " Generalized eigenvalue are solved by Lanczos, which has high precision and high operation speed. It allow poor quality solid element in finite element model. It is always getting the more modes of the big models with Symmetrical characteristic.[3] The Solution of Boundary Conditions as the Bearing Constraint is Rigid Constraint. When the bearing are regarded as rigid constraint, the full constraints are exerted on the nodes of eccentric shaft.After exerting the constraints of eccentric shaft is shown in Fig.2. The low natural mode of vibration has more influence than high natural mode of vibration on structure, and the low natural mode of vibration play decision role on dynamic characteristics, so solve and extract the first six natural frequency and model shape, which are shown in Table.1 and Fig.3. The critical speed: fn 60= and f is frequency. Afterwards the natural frequency are converted into the critical speed which is shown in Table.1. Table.1 The Frequency and Critical Speed of Every Modal Order 1 2 3 4 5 6 Frequency(Hz) Speed(r/min) 368"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003336_20110828-6-it-1002.03057-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003336_20110828-6-it-1002.03057-Figure2-1.png",
        "caption": "Fig. 2. Two rotating inertias interacting by friction (a) and equivalent system obtained when z2 = 0 (b).",
        "texts": [
            " Aim of this paper is to extend the use of the model of Fig. 1 to this class of systems. In this section, the model illustrated in previous section is adapted to correctly simulate the behavior of systems composed by a number of objects that interact in pairs by means of frictional interfaces. Since, the goal of this research concerns the simulation of clutches used in the automotive field, rotative systems are taken into account, but the same considerations hold true for translating systems. Given the system of Fig. 2(a), composed by two rotating bodies with inertia J1 and J2 respectively, described by the dynamic model { J1\u03c9\u03071 = F1 \u2212 \u03c412(\u03c91 \u2212 \u03c92) J2\u03c9\u03072 = F2 + \u03c412(\u03c91 \u2212 \u03c92) (5) where \u03c9i and Fi are the angular velocity and the external torque related to the i-th mass and \u03c4i,i+1(\u00b7) denotes the friction at the interface between the i-th and the (i + 1)- th object, it is possible to obtain the same formulation as in (3), by means of a proper congruent state space transformation. Firstly, it is convenient to rewrite the system (5) in a matrix form as J\u2126\u0307 = F \u2212D T \u03c4 (D\u2126) (6) with J = [ J1 0 0 J2 ] , \u2126 = [ \u03c91 \u03c92 ] , F = [ F1 F2 ] , D = [1 \u2212 1] and \u03c4 = \u03c412",
            " (2004) among many others), but it is worth noticing that the approach based on the congruence transformation T provides a systematic procedure for the computation of the friction at zero velocity. As a matter of fact, it is sufficient to solve the algebraic equation obtained from (9) by assuming z\u03072 = 0. This result is quite intuitive since the static friction opposes the motion and, therefore, counteracts all the external torques in order to guarantee that the relative velocity remains zero. When z2 = 0, the system is completely described by state variable z1 and by the equation (8); as shown in Fig. 2(b) it behaves like a unique inertia J1 + J2, subject to the resultant of all the external torques. Note that, by substituting \u03c41,2 = J2F1\u2212J1F2 J1+J2 in both the equations of the original system (5) we just obtain (8). Given the system composed by three rotating masses shown in Fig. 3, that are subject to external torques and friction between the contacting masses, as described by the system of differential equations    J1\u03c9\u03071 = F1 \u2212 \u03c412(\u03c91 \u2212 \u03c92) J2\u03c9\u03072 = F2 + \u03c412(\u03c91 \u2212 \u03c92)\u2212 \u03c423(\u03c92 \u2212 \u03c93) J3\u03c9\u03073 = F3 + \u03c423(\u03c92 \u2212 \u03c93) (10) one may decouple the main dynamics that does not depends on the internal frictional torques from the other relative dynamics"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.37-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.37-1.png",
        "caption": "Figure 12.37 Path of the top symmetry axis on a sphere when \u03d5\u0307 (t) is changing between two positive extreme values.",
        "texts": [
            "630) cos \u03b81 = u1 cos \u03b82 = u2 (12.631) Therefore, \u03d5\u0307, \u03b8 , and \u03c8\u0307 are elliptic functions of time. The period of these functions is half the period of sn\u03c4 . The superposition of periodic changes of \u03b8(t) onto a precession about the Z-axis with a periodically changing angular velocity \u03d5\u0307 (t) generates a wavy motion of the z-axis. It can be visualized on a sphere with the center at the fixed point X Z \u03c8 \u03d5 Y \u03b8 z The general paths of motion of the symmetry axis of the top are illustrated in Figures 12.37\u201312.39. Figure 12.37 depicts the situation in which \u03d5\u0307 (t) is changing between two positive extreme values. Figure 12.38 shows the path of the z-axis when \u03d5\u0307 (t) is changing between negative and positive values. Figure 12.39 depicts the situation in which \u03d5\u0307 (t) is changing between zero and a positive maximum. The periodic dance motion of the top with \u03b8 (t) which is superimposed on the precessional motion is called nutation. Example 756 Special Cases of Top Dynamics Some special cases of top dynamics have simpler equations of motion with simpler interpretations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003785_icnmm2013-73187-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003785_icnmm2013-73187-Figure1-1.png",
        "caption": "Fig. 1 Schematic of experimental apparatus",
        "texts": [
            " Then, in the case of low-volume manufacturing, a spray nozzle method can be used but it is an inefficient production method. Recently, extremely small beads have been produced using T and Y shaped microchannels(10). An orifice method(11)(12) is able to produce uniform microbeads, however, these micro devices are quite expensive and prevention of clogging around the orifice in the microchannel is required. We focus our attention on the nozzle method and investigated a simpler method of manufacturing uniform microbeads. By improving the previous nozzle method, we propose a new method. Figure 1 shows a schematic diagram of the experimental device to produce microbeads. An aqueous solution of sodium alginate containing a ferrofluid squirts out of a nozzle by a syringe. Droplets fall into an aqueous calcium chloride solution and form microbeads. The microbead diameter could be adjusted by changing the liquid level in the beaker and the flow rate of the aqueous sodium alginate solution as shown Fig. 1. The flow rate was controlled by the syringe from 0.001ml/s to 0.07 ml/s. In the present study, the proper distance between the liquid surface and nozzle was found to be 0.56 mm. Figure 2 shows the microbeads produced by the present method. Since the droplet formed at the nozzle kept getting bigger until it fell down under its own weight as shown in Fig. 2(a), this method of droplet formation at the nozzle was not suitable for making small beads. On the other hand, Figure 2(b) shows the microbeads formed after a free fall and hitting the liquid surface at which the droplet was torn apart by an adsorption force"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001813_9781118188347.ch7-Figure7.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001813_9781118188347.ch7-Figure7.9-1.png",
        "caption": "Figure 7.9 The structure of a Hall position sensor.",
        "texts": [
            " Position sensors in BLDCmotors are used to detect the relative position of the rotor magnet and provide the correct commutation information for the logic switching circuit, namely transforming the position signals of rotor magnet to electric signals and then making stator windings commutate properly. The commonly used position sensors mainly fall into electromagnetic, photoelectric and magnetic types. A Hall position sensor, as one kind of magnetic-type sensors, is applied extensively for its simple structure and low cost. A Hall position sensor, shown in Figure 7.9, is constituted by a Hall integrated circuit fixed on stator and sensor rotor fixed on the main rotor in most BLDC motors. The sensor indicates the main rotor\u2019s position since its rotor is rotating with the motor rotor synchronously. Several Hall integrated circuits are fixed on the motor\u2019s stator at equal intervals and the sensor will produce a group of jumping signals when the main rotor passes by a pair of magnets. The more pole pairs of the main rotor, the more jumping signals are generated within 360 mechanical angles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003848_imece2012-87624-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003848_imece2012-87624-Figure4-1.png",
        "caption": "FIGURE 4. DEFORMATION OF BODY Ki",
        "texts": [
            " The deformation of the body is described by the nqi shape functions [15] of the body Ki given by \u03a6i(R), where R describes the position of the point R within the reference frame of the body Ki. The displacement of a material point R within the the body Ki is given by \u2206ui = \u03a6i l(R i)qil , where qil is the coordinate of the l-th shape function of the body i. A frame attached to the material point R is subjected to the rotation A\u0303(R) with respect to the body frame ~ei, which can be described by e.g. the Euler angles \u03bd(Ri) given by \u03bd(Ri) = \u03a8i(Ri)qi. The local rotation angles \u03a8i(Ri) are part of the shape function description. The deformation of the body Ki is depicted in Figure 4. The state of the body Ki can therefore be described by the position of the reference frame xI i \u2208 Rn i p+n i \u03b1 , with nip as the dimension of the space and ni\u03b1 gives the dimension of the needed to store the orientation vector, and the current coordinates of the shape function qi, xI i = [ pi Ai ] and zI i = [ xI i qi ] , 1in case of subindices the Einsteinian notation is used with appropriate dimension, superindices are exempted, e.g. aT b = akbk = \u2211 k akbk 2 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002789_amr.816-817.289-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002789_amr.816-817.289-Figure1-1.png",
        "caption": "Fig. 1 Geometry of the hybrid bearing (a) Cross section view (b) Developed view of the bearing surface.",
        "texts": [
            " A two-dimension finite element method considering turbulence effect and viscosity change with water film temperature has been developed to solve the Reynolds equation to obtain different bearing performance. 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: 142.103.160.110, University of British Columbia, Kelowna, Canada-12/07/15,18:57:37) The geometry structure of the water-lubricated hybrid bearing and its coordinate system rotor are shown in Fig. 1.In the inner surface of the bearing, there are four stepped recess, four restrictor inlets and eight cooling inlets. The ring grooves on the water seal surface are connected with the deep recess, which form eight T-shaped grooves. The T-shaped grooves connect all the restrictor inlets and cooling orifices, which can limit temperature, rise and prevent film rupture on the bearing surface. Shallow recess surrounded by the T-shaped grooves can strengthen the dynamic pressure to improve the load carrying capacity and stiffness"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure8-1.png",
        "caption": "Figure 8 \u2212 Scheme of the test rig",
        "texts": [
            " Fitting the castor free oscillation with a damped harmonic function, the oscillation frequency and the damping ratio were deduced. Applying this procedure for several stiffness values, the diagram reported in figure 7 was obtained; it reports the damping ratio versus the lateral stiffness and shows the threshold stiffness value for which the system changes from a stable to an unstable behavior. 3 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/75795/ on 04/09/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use The test rig (Fig. 8) consists of a castor, derived from a scooter front assembly, joined to a rigid steel frame by means of a support that allows the castor to vertically translate and rotate around its steering axis. Furthermore, the support allows to adjust the rake angle. A vertical load can be added to the castor, by means of masses joined to the support via a wire rope and a system of pulleys (Fig. 9). Castor rotation is counteracted by a 18 position (click) adjustable steering damper. The castor wheel rolls on a flat track, made up of a composite material belt, wrapped on two rolls one of which is driven by an electric motor [9]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-FigureB-3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-FigureB-3-1.png",
        "caption": "Figure B-3. Output Angle per Tooth Hit on the Ring Gear",
        "texts": [
            " For planet gears (Figure B-1), they will get hit once when \u03b8C + (-\u03b8P) = 360\u25cb (B1) From Table 1, the following rotational relationship can be derived as: R S P S R S C P N N N N N N \u2212 = \u03b8 \u03b8 (B2) After substituting Equation (B2) into (B1), the output angle per tooth hit on the planet gear can be derived as: \u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b = R P C N N 1 360o\u03b8 (B3) For the sun gear (Figure B-2), it will get hit once when \u03b8S - \u03b8C = 360\u25cb/N (B4) From Table 1, the following rotational relationship can be derived as: R S R S C S N N N N 1+ = \u03b8 \u03b8 (B5) After substituting Equation (B5) into (B4), the output angle per tooth hit on the sun gear can be derived as: Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 10 Copyright \u00a9 2011 by ASME \u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b = R S C N N N o360\u03b8 (B6) For the ring gear (Figure B-3), it will get hit once when NC o360 =\u03b8 (B7) Table 1. Rotational Relationships of a Simple Planetary Gear Set Description Input Sun Planet Gear Ring Gear Carrier Output 1 Rotate the sun once while holding carrier +1 -NS NP -NS NR 0 2 System is fixed as a whole and rotated NS NR NS NR NS NR NS NR NS NR \u03b8S \u03b8P \u03b8R \u03b8C 3 1 + 2 NS NR + 1 NS NR - NS NP 0 NS NR Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002774_mmar.2012.6347894-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002774_mmar.2012.6347894-Figure2-1.png",
        "caption": "Fig. 2. Robot with virtual redundant axis q7",
        "texts": [
            " However, this leads to a mix of errors and to jerky motions at the end-effector, caused by switching between singularity avoidance and normal operation. In some settings of base-motion compensation, for example if the base-motion compensation is used for sensitive cargo at the end-effector, such mixed errors and jerky motions may be disadvantageous, leading to damage or injuries. For this reason, we propose a new approach in which all errors induced by singularity avoidance can be concentrated in one individual direction. The basic idea is to introduce a virtual redundant axis at the end-effector\u2019s end (Fig. 2) and to resolve the virtual redundancy problem such that all tracking errors are \u2019absorbed\u2019 by this virtual redundant axis. Then, the robot will rotate about this virtual redundant axis when approaching singularity, reducing errors in direction of all other task-space directions ideally to zero. With the virtual redundant axis, the direct kinematics becomes rAE = \u03a6\u0304(q\u0304) , (8) where vector q\u0304 = [q1 , q2 , q3 , q4 , q5 , q6 , q7] T collects the robot axes as well as the virtual redundant axis q7, and KE is the end-effector frame"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002357_ecce.2013.6647276-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002357_ecce.2013.6647276-Figure6-1.png",
        "caption": "Fig. 6. 4-pole/12-slot winding configuration (only the U-phase is shown).",
        "texts": [
            "5\u00b0 are 24 mNm and 19 mNm, respectively, and these values become smaller than that with 8-pole/9-slot configuration, because the number of stator slots is changed from 9 to 12, and thus, the number of coil turns around each tooth is decreased. However, the torque ripples when \u03b8m = 0\u00b0 and \u03b8m = 7.5\u00b0 are decreased to 4 mNm and 5 mNm, respectively. It turns out that the 8-pole/12-slot configuration contributes to a reduction in torque ripple. C. 4-pole/12-slot Configuration For further reduction in torque ripple, the 4-pole/12-slot configuration was investigated, and the results are reported in this section. Figure 6 shows the analysis models with different winding arrangements. Only the U-phase winding is shown in the figures. The V- and W- phases are arranged to have a shift from the U-phase electrically by 120\u00b0 and 240\u00b0, respectively. In model 1, the U-phase consists of 4 main coils with N turns. However, model 2 has sub-coils with 0.5N turns located on both sides of the main coil. The sub-coils and the main-coils in each phase are connected in series. Figure 7 (a) and (b) show the average and ripple of the calculated torque, respectively",
            " This waveform distortion may influence the torque ripple. The variation of the magnetic flux density depending on the magnetized rotational angle \u03b8m still remains in both models of the 4-pole/12-slot configuration. D. 2-pole/12-slot Configuration To reduce the influence of the magnetized rotational angle on the torque ripple variation, a 2-pole/12-slot configuration was investigated and is reported in this section. When models 1 and 2 were applied to the 2-pole motor, the phase current direction on one side shown in Fig. 6 was reversed. In addition to the aforementioned models 1 and 2, model 3 is discussed as shown in Fig. 10. This winding arrangement has previously been proposed by the authors with the intent of reducing interference and variation of the magnetic suspension force in a bearingless motor [7]. This winding arrangement contributes to reduction in the harmonic components of the stator\u2019s MMF distribution. Additional short-pitch sub-coils were wound around the stator teeth on both sides of the main coil"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002995_scis-isis.2012.6505293-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002995_scis-isis.2012.6505293-Figure2-1.png",
        "caption": "Fig. 2. Interpolation using Gouraud shading algorithm.",
        "texts": [
            " 2) Linear Interpolation Algorithm: The simple linear interpolation algorithm is applied to calculate the agent\u2019s move position. This algorithm is same as Gouraud shading algorithm [7]. Gouraud shading algorithm is a method used in the computer graphics domain to simulate the differing effects of light and color across the surface of an object. Because triangles generated by Delaunay Triangulation are bordered on adjacent triangles by their edges, it is guaranteed that the output value is continuous in the covered region. Fig. 2 shows the process of Gouraud shading algorithm. The output values from vertices Pa, Pb and Pc are O(Pa), O(Pb) and O(Pc) respectively. Now, we want to calculate O(B) , the output value of the point B contained by the triangle PaPbPc. The algorithm is as follows: 1) Calculates I , the intersection point of the segment PbPc and the line PaB. 2) The output value at I , O(I), is calculated as: O(I) = O(Pb) + (O(Pc) \u2212 O(Pb)) m1 m1 + n1 where |\u2212\u2212\u2192PbI| = m1 and |\u2212\u2212\u2192PcI| = n1. 3) O(B) is calculated as: O(B) = O(Pa) + (O(I) \u2212 O(Pa)) m2 m2 + n2 where |\u2212\u2212\u2192PaB| = m2 and |\u2212\u2192BI| = n2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003500_iccas.2013.6704046-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003500_iccas.2013.6704046-Figure5-1.png",
        "caption": "Fig. 5 Mathmatical model of the developed robot",
        "texts": [
            " In this time, no force acts on the proposed robot's actuator, except for a friction force. Secondly, we consider the time the arm frame moves in the different as the robot's traveling direction, as shown in Fig. 4 (b). At this moment, the passive wheel keeps pointing same di rection by friction force, and then the force by the rotary spring acts on the wheel. Thus, the passive wheel turns in original direction by this force. As a result, the proposed robot can move forward. We consider about the time that the driving torque by a servo motor acts on the passive joint, as shown in Fig. 5 . nand t represent the normal line direction and the tangent line direction for the wheel rolling direction, respectively. La and Lw indicate the arm frame length and the passive wheel length, respectively. The moment by the servo mo tor is defined as lvI. Moreover, the force generated by a spring acts on the passive wheel toward n direction, and then the spring constant is defined as k. Rt and Rn rep resent the reaction forces toward t and n, respectively. Defining the total forces as Fn (normal direction) and Ft (tangental direction), respectively, we have the following equations: M Ft = sin e - Rt, La + Lw cose M Fn = L L e cos e + ke - Rn, a + w cos (I) (2) where e is angular difference between the tangental di rection and the direction along the arm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure9-1.png",
        "caption": "Figure 9 \u2212 Castor on the belt rig",
        "texts": [
            "url=/data/conferences/asmep/75795/ on 04/09/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use The test rig (Fig. 8) consists of a castor, derived from a scooter front assembly, joined to a rigid steel frame by means of a support that allows the castor to vertically translate and rotate around its steering axis. Furthermore, the support allows to adjust the rake angle. A vertical load can be added to the castor, by means of masses joined to the support via a wire rope and a system of pulleys (Fig. 9). Castor rotation is counteracted by a 18 position (click) adjustable steering damper. The castor wheel rolls on a flat track, made up of a composite material belt, wrapped on two rolls one of which is driven by an electric motor [9]. The test rig allows hence to set: a) the belt speed; b) the rake angle; c) the vertical load; d) the fork suspension extension; e) the steer damping. Since the force exerted by the steering damper depends almost linearly on the steer rotating speed, the damping value for each of the first nine damper positions (clicks) was estimated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002724_s1052618811020142-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002724_s1052618811020142-Figure4-1.png",
        "caption": "Fig. 4. The different structures of rotary tables with two degrees of freedom.",
        "texts": [
            " 3): \u2013 \u03b4x(xrated) = xreal \u2013 xrated deviation during positioning along X coordinate (EXX, line 5); \u2013 deviation from linearity \u03b4y(xrated) that takes place in XY plane (EYX, line 7); \u2013 deviation from linearity \u03b4z(xrated) that takes place in XZ plane (EZX, line 3); \u2013 \u03b1x(xrated) rotation around X axis (EAX, line 4); \u2013 \u03b1y(xrated) rotation around Y axis (EBX, line 8); \u2013 \u03b1z(xrated) rotation around Z axis (ECX, line 2); \u2013 coordinates of measurement lines of \u03b4x(xrated), \u03b4y(xrated), and \u03b4z(xrated) linear deviations by assuming that the pole of motion is in the point of their intersection. Revolute pair. The same approach for describing the primary deviations can be used for the drive according to the angular coordinate. Figure 4 depicts the different structures of rotary tables with two degrees of freedom. Figure 4a shows the rotary table of the cantilever structure for the MC 300 machine, and Fig. 4b presents the rotary table of globe structure for the MC 700 machine (1 is the faceplate, and 2 is the faceplate body). The faceplate is intended for rotating the processing component according to the C coordinate. The faceplate body is rotated (depending on the table structure) either around the Y axis (angular coordinate B) (Fig. 4a) or around the X axis (angular coordinate A) (Fig. 4b). The block diagram of the faceplate rotary mechanism, according to two independent degrees of freedom, is the same for both structures of the rotary table and is presented in Fig. 1b. 110 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 40 No. 2 2011 SERKOV Figure 5 depicts the block diagram of the drive, which is controlled according to the rotary table\u2013face plate program. Here, there are the same elements that are used in the drive, according to the linear coor dinate, but the rotary table is used instead of the slide, and the angle of rotation sensor is used instead of the linear motion sensor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003854_20120905-3-hr-2030.00044-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003854_20120905-3-hr-2030.00044-Figure11-1.png",
        "caption": "Fig. 11. Concept of the virtual leg",
        "texts": [
            " We thus make a hypothesis that push-off due to the ankle elasticity generates torque that is similar to VAT. To confirm this hypothesis, we transform the push-off force into the torque around the contact point of the stance leg. We define that L1 (m) is a length of the virtual stance leg, L2 (m) is a length of the virtual swing leg, \u03b31 (rad) is an angle of the virtual stance leg, \u03b32 (rad) is an angle of the virtual swing leg, \u03b1 = \u03b84 \u2212 \u03b83 (rad) is the relative angle between the shank and foot of the swing leg and F (\u03b1) (N) is a push-off force as shown in Fig. 11. For simplicity, we then assume l1 \u2248 0 (m), l2 \u2248 0 (m), l3 \u2248 1.0 (m), m1 \u2248 0 (kg), m2 \u2248 0 (kg), m3 \u2248 5.0 (kg), D \u2248 0 (Nm/(rad/s)) and \u03b2 \u2248 0 (Nm/(rad/s2)). When these assumptions are almost satisfied, we can notice L1 = L2 = L and \u03b31 = \u03b32 = \u03b3 as shown in Fig. 12. We can, moreover, infer that a linear relationship exists between \u03b1 and F (\u03b1) since the push-off force substantially depends on the restoring torque due to the ankle elasticity of the swing leg. We thus assume that the push-off force is F (\u03b1) = \u2212\u03b5\u03b1 and the direction of F (\u03b1) is equal to the direction from the hip of the swing leg to the tip of the swing leg"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003610_imece2013-63166-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003610_imece2013-63166-Figure1-1.png",
        "caption": "FIGURE 1: SKETCH OF THE MANIPULATOR. (a) VERTICAL; (b) HORIZONTAL",
        "texts": [
            " In this paper, the rotation of the manipulator about an axis contained in the (Pu,v) plane of the mobile platform is studied for two other transformation combination possibilities. The study of the rotation order for these combinations shows, that the parasitic motion expressions are different for each transformation order and that the value of parasitic motion is different too. This leads to different power and energy in the actuators. Taking this into account, energy optimization could be done. Figure 1 shows the mechanism analysed in this work in horizontal and vertical configurations. It consists of a mobile and a fixed platforms joined by three identical linkages, equally spaced at 2.0944 rad (120\u25e6). The manipulator has three actuated prismatic joints (P), whose lower ends are attached to the base by non-actuated revolute joints (R). The upper end of the prismatic joints are joined to the moving platform by spherical joints (S). Applying the mobility criterion presented in [4], the mechanism is a 3-DOF manipulator. The 3 DOF achievable are a vertical translation and two rotations about two perpendicular axes intersecting at the mobile platform center. As shown in Fig. 1, a base coordinate frame, (OX ,Y,Z), is fixed at the center of the base platform with the Zaxis vertical and the X-axis pointing towards the point B1 in the horizontal configuration and towards the point C1 in the vertical one. In a similar way, the moving coordinate frame, (Pu,v,w), is defined with the w-axis normal to the platform plane and the uaxis pointing towards the point A1. According to this, the 3-DOF are a translation along the Z-axis and two rotations about the Xand Y- axes (\u03c8 and \u03b8 , respectively)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.45-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.45-1.png",
        "caption": "Fig. 2.45 4PaPRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\P\\kR\\Pa",
        "texts": [
            "21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig. 2.49) R\\Pa||Pa\\kR (Fig. 2.22j) The first and the last revolute joints of the four limbs are parallel 38. 4RPaPaR (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002413_1350650112462940-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002413_1350650112462940-Figure1-1.png",
        "caption": "Figure 1. Bearing schematics (not to scale, for illustration purposes only).",
        "texts": [
            " Let F1 and F2 symbolize the high-viscosity and the low-viscosity lubricant, respectively, in a journal bearing. Assume that the respective viscosities 1 and 2 of the two liquids F1 and F2 are significantly different. Liquid F1, the high-viscosity lubricant, is made available to the shaft surface, which drags it into the clearance space S1 of the bearing, thus creating a \u2018thick\u2019 film of F1. In contrast, lubricant F2, the low-viscosity component, is introduced adjacent to the bearing in clearance space S2 and is dragged along by the floating-ring. In this manner, a \u2018thin\u2019 film ofF2 is created (Figure 1). The resulting composite film {F1\u00feF2} adequately separates the surfaces, while, in accordance with the encapsulation instability principle we referred to above, deformation work localizes to component F2. We show that this new version of the CFB performs not unlike the first one; specifically, its application leads to substantial reduction of viscous losses in all cases but in various degrees, depending on system construct. To keep the numerics simple, we consider an approximate model of journal-bearing lubrication, long bearing with Gumbel\u2019s boundary condition",
            " In addition, we make use of the symbol for the ratio of surface velocities, and the approximation \u00bc ub uj \u00bc rbnb rjnj nb nj \u00f03\u00de The variables of the problem are normalized according to x \u00bc r ; z \u00bc lZ; h=c \u00bc 1\u00fe \" cos \u00f0 \u00de p \u00bc nj 1\u00fe \u00f0 \u00de r c 2 P The shear forces fj and fb acting on the two bounding surfaces are obtained from equation (2) by integration, and the friction variable, evaluated on the journal and on the bearing, respectively, has values r c f w j,b \u00bc \" sin \u2019 2 \u00fe 1 \u00f0 \u00de 1\u00fe \u00f0 \u00de 2Sffiffiffiffiffiffiffiffiffiffiffiffi 1 \"2 p \u00f04\u00de Here S is the Sommerfeld number, defined by S \u00bc nj 1\u00fe \u00f0 \u00de w=2rjl rj c 2 \u00f05\u00de where w is the external load on the shaft and l is the length of the bearing. The factor 1\u00fe \u00f0 \u00de in equation (5) takes into account that for a journal bearing in which both shaft and bearing are rotating, the rotational speeds are additive. The floating ring divides the clearance space S \u00bc S1 \u00fe S2f g into two non-connecting regions, S1 adjacent to the shaft and bounded by shaft and floating-ring, and S2 bounded by floating-ring and bearing. In what follows, we designate r1 and r2 to represent the inner radius and outer radius of the floating-ring, respectively (Figure 1). The above formulas assume distinctly different values in the two regions. c \u00bc c1; nring=nj; \u00bc 1; rj r1; \" \u00bc \"1 S \u00bc S1 \u00f06\u00de The definition of the Sommerfeld number S1 follows from equation (5) S1 \u00bc 1nj 1\u00fe \u00f0 \u00de w=2r1l\u00f0 \u00de r1 c1 2 \u00f07\u00de The friction variable, calculated for the inner surface of the floating ring, is rf cw 1 \u00bc 1 1\u00fe 2S1ffiffiffiffiffiffiffiffiffiffiffiffi 1 \"21 q \"1 sin \u20191 2 \u00f08\u00de In contrast, the friction variable on the journal is obtained from rf cw j \u00bc 1 1\u00fe 2S1ffiffiffiffiffiffiffiffiffiffiffiffi 1 \"21 q \u00fe \"1 sin\u20191 2 \u00f09\u00de Region S2 c \u00bc c2; nb nring \u00bc 0; \u00bc 2; rj \u00bc r2; \" \u00bc \"2 S \u00bc S2 \u00f010\u00de The Sommerfeld number S2 has the definition S2 \u00bc 2nring w=2r2l\u00f0 \u00de r2 c2 2 \u00f011\u00de Moreover, the friction variable on the outer surface of the floating ring is rf cw 2 \u00bc 2S2ffiffiffiffiffiffiffiffiffiffiffiffi 1 \"22 q \u00fe \"2 sin \u20192 2 \u00f012\u00de at Library - Periodicals Dept on November 15, 2014pij",
            " Moreover, for identical bearings and under the conditions of this paper, the normalized bearing-performance parameters, the Sommerfeld number in particular, depend on the eccentricity ratio alone. We have a choice, therefore, to display bearing-performance results against the eccentricity ratio \"1 of the inner bearing, or any of the Sommerfeld numbers defined above. On equipping the bearing with a second film, new input parameters, the viscosity ratio \u00bc 2= 1, the ratio of the ring outer to inner radii R \u00bc r2=r1, and the clearance ratio C \u00bc c2=c1, appear (Figure 1). Figures 2 to 6 investigate the individual influences of these parameters, changing one parameter at a time while keeping the others constant. Obviously, the combined effect of changing all three parameters simultaneously will not equal the sum of three partial changes; the problem is nonlinear in the parameters. Figure 2 portrays the changes in the eccentricity ratio of the outer bearing \"2, the load coefficient Cw, and the power coefficient C+, for three values of the clearance ratio C \u00bc 0:5, 0:3 and 0:1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure8-1.png",
        "caption": "Figure 8 Multi-directionally loaded raster orientation tensile specimen (see online version for colours)",
        "texts": [
            " If the stresses in the laminate axes are denoted by x, y and xy, then these are related to the stresses referred to the material axes by the usual transformation equations: 2 2 1 2 2 2 2 2 12 2 2 x y xy c s cs s c cs cs cs c s (1) where c denotes cos and s denotes sin . Also, the strains in the material axes are related to those in the laminate axes, namely x, y and xy, by what is essentially the strain transformation: 2 2 1 2 2 2 2 2 12 2 2 x y xy c s cs s c cs cs cs c s (2) Consider a similar approach using FDM, whereas the tensile specimen highlighted in Figure 5 is shown from a top view perspective in Figure 8. Raster orientations are additively constructed in the FDM process and specified as toolpaths in the Insight software. Therefore, a natural expansion of thought would be to consider that each ply in composite laminate theory may be thought of as a raster pattern layer in the FDM process and by specifying each layers orientation to 0\u00b0 each ply or layer, the specimen may be defined to serve as a function of loading for multi-directionally loaded parts. Each specimen offers unique orientations to correlate to the laminate theory fundamental principles listed above. By tailoring several orientations within the same tensile specimen in a circular pattern, the specimen shown in Figure 8 represents a multi-directionally loaded part. From a testing perspective, each multi-directionally loaded specimen group would yield the following quantity of tensile bars listed in Table 2. Multiple groups may be built within the 400MC build chamber to allow a statistically significant population to be generated in the same build. By testing a number of these groupings from the same build, one may gain raster orientation direction mechanical property data relative to the Z-axis and a thorough knowledge capture may be attained for the FDM process"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.70-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.70-1.png",
        "caption": "Figure 2.70. Wound rotor induction machine: (a) rotor with slip rings; (b) slip-ring motor",
        "texts": [
            " The first of these is simpler in construction and is more frequently used. The winding of such a rotor is in form of a cylindrical cage (the so-called \u201csquirrel cage\u201d) of copper or aluminum bars, short-circuited around the ends by two rings. The bars of this winding are placed without insulation in the slots of the rotor (Figure 2.69a). Another method in use is to pour molten aluminum into the slots of the rotor. A slip-ring rotor, also known as a phase-wound rotor, has a winding made of insulated wire (Figure 2.70a) and in most cases in three-phase wye connected. The free ends of this winding are brought out to the slip rings on the rotor shaft. Brushes bear on the slip rings and connect the rotor winding to a three-phase rheostat (Figure 2.70b). This system permits the resistance of the rotor circuits to be varied. This is very important in starting the motor. An asynchronous machine is similar to a transformer in the sense that the power is transferred from the stator (primary) to the rotor (secondary) winding only by mutual induction. For this reason an asynchronous machine is often called an induction machine. The induction machine operation is due to the fact that the ratio of the rotor angular speed and the network frequency varies with the motor load and characteristics of supply"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003110_s1068798x11060153-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003110_s1068798x11060153-Figure2-1.png",
        "caption": "Fig. 2. Body roll of a turning tricycle.",
        "texts": [
            " 1, O, O ' are the centers of rotation of tricy cles with rigid and elastic wheels; R is the distance from the center of rotation to the tricycle\u2019s longitudi nal symmetry axis; B, L are the track and base of the tricycle (L = a + b, where a and b are the distances from the center of mass to the tricycle\u2019s front and rear axes); F, Fx, Fy are the centrifugal force and its longi tudinal and transverse components; Y1, Y2 are the lat eral (transverse) reactions of the supporting surface; \u03b8 is the angle of rotation of the controlled wheel; \u03b41, \u03b42 are the angles of lateral displacement of the tricycle\u2019s axes; \u03c1 is the radius of rotation of the tricycle\u2019s center of mass C; \u03b3 is the inclination of \u03c1 to the normal to the tricycle\u2019s longitudinal axis; Fyt is the transverse com ponent of the centrifugal inertial force relative to the tricycle\u2019s tipping axis (Fig. 1b). Consider the motion of a tricycle on a horizontal supporting surface in rotation, when a component of the centrifugal inertial force that is transverse to the tipping axis acts on the tricycle\u2019s body (Fig. 2). In Fig. 2, we show the normal reactions (Z1, Zex, Zin) and transverse reactions (Y1, Yex, Yin) of the road acting on the tricycle\u2019s single wheel and the wheels external and internal to the center of rotation. Superscripts s and n denote masses of the tricycle that are and are not subject to the springs. In Fig. 2, are components of the centrifugal inertial forces that act on the masses subject to and not subject to the springs as the tricycle turns and that are trans verse to the tipping axis; G = mg is the weight of the tri cycle; G s, Gn are the weights of the masses subject to and not subject to the springs; m is the mass of the tri cycle; ms, mn are the masses subject to and not subject Fyt n Fyt s , Fyt n to the springs; g is the acceleration due to gravity; \u03bb is the roll angle of the body (the masses subject to the springs); h s, hn = r are the heights of the centers of mass of the sections of the tricycle subject to and not subject to the springs; r is the wheel\u2019s rolling radius; h\u03bb is the distance at which the roll acts; C s, Cn are the centers of the tricycle\u2019s masses subject to and not sub ject to the springs in linear motion; \u0394 is the lateral dis placement of the center of the elastic wheel\u2019s contact spot. (We assume that \u03941 = \u0394ex = \u0394in = \u0394.) The distance (relative to the tipping axis) at which the tractional force of the body (the mass subject to the springs) is applied is (Fig. 2) (1) where \u03b1 is determined from the equation tan\u03b1 = B/(2L). Analogously, we may determine the distance (rela tive to the tipping axis) at which the tractional force of the mass that is not subject to the springs is applied (2) ns B/2 \u0394\u2013 h\u03bb \u03bbsin\u2013( )acos\u03b1 L ,= nn B/2 \u0394\u2013( )a \u03b1cos L .= DOI: 10.3103/S1068798X11060153 RUSSIAN ENGINEERING RESEARCH Vol. 31 No. 6 2011 BODY ROLL AND THE STABILITY OF A TRICYCLE WHEN TURNING 583 Comparing Eqs. (1) and (2), we find that (3) The transverse component of the tricycle\u2019s centrif ugal force relative to the tipping axis is (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002407_amr.189-193.1409-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002407_amr.189-193.1409-Figure4-1.png",
        "caption": "Fig. 4. The decoupled spherical parallel mechanism",
        "texts": [
            " Since the limb constraint system consist one constraint force 1f , the twists of limb can be obtained: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= , 1 4 4 4[0,0,0; , , ]P d e f= , 2 5 5 5[0,0,0; , , ]P d e f= The new revolute joints 4 5,R R can be obtained by the linear combination of revolute joint and prismatic joint, so the twist system of limb could become: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= 4 1 1 1 4 4 4[ , , ; , , ]R a b c d e f= , 5 2 2 2 5 5 5[ , , ; , , ]R a b c d e f= Therefore, the UPU limb can be obtained to afford the constraint force 1f , as shown in Fig. 2. In the similar method, the SPS limb can be obtained to afford the constraint no force, as shown in Fig. 3. Finally, The symmetrical decoupled SPM (Fig. 4-a) and non-symmetrical decoupled SPM(Fig. 4-b) can be obtained by assembling the limbs synthesized perpendicularly. Since each UPU limb affords one constraints f1, the 3 UPU\u2212 parallel mechanism (Fig. 4-a) and 2SPS RU\u2212 parallel mechanism (Fig. 4-b) is non-over-constrained. The principle of decoupled motion of UPU RU SPS\u2212 \u2212 SPM are as following: when the active prismatic joint of UPU chain works, the moving-platform of spherical parallel mechanism rotates around axis X (actuated by the motor 1), while the actuators of two kinematic chains holds as the kinematic chain has the kinematic joint along axis X. Similarly, when the actuators of RU and SPS kinematic chains work, the moving-platform of spherical parallel mechanism rotate around Y and Z axis respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000295_032075-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000295_032075-Figure1-1.png",
        "caption": "Figure 1. The scheme of the aggregate consisting of the tractor and of the mounted tillage unit taking into account possible displacements and active forces in the longitudinal-vertical plane where the coordinates of the centers of gravity is marked with: C1 \u2013 the energy machine; C2 \u2013 supporting wheel; C3 \u2013 frame of the unit; C4 and C5 \u2013 the first and second rows of the flat-cutting hoes; C6 \u2013 disc sections.",
        "texts": [
            " The theoretical research were conducted the stability of the unit in the longitudinal-vertical plane when performing tillage. Considered are two options for placement of disc sections in the unit: in the ESDCA 2021 IOP Conf. Series: Earth and Environmental Science 723 (2021) 032075 IOP Publishing doi:10.1088/1755-1315/723/3/032075 first embodiment the disc section located behind of flat-cutting hoes or cultivator hoes, in the second variant \u2013 placed front of flat-cutting paws or cultivator paws. Consider the force interaction of the machine-tractor aggregate with the supporting surface of soil (figure 1). To do this we will use the results of research by A. B. Lurie and other scientists [10-12]. The system \"tractor + mounted tillage unit\" is considered as a two-mass system: the mass m\u0422R is located in the C1 center mass of the tractor and a mass mUN of the tillage unit consisting to the sum of the masses of frame, of supporting wheel, of flat-cutting hoes or of cultivator hoes and of disk sections is located in the center of mass \u0421UN or are applied to the axis of rotation of the unit. We take two coordinate systems to describe the movement of a machine-tractor aggregate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000295_032075-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000295_032075-Figure2-1.png",
        "caption": "Figure 2. Change in the period of free vibrations of the mounted tillage unit into subjection on the disposition of the operating bodies: 1 \u2013 the flat-cutting paws are located in front of the disc sections; 2 \u2013 the disc section are located in front of the flat-cutting paws; 3 \u2013 the cultivation paws are located in front of the disk sections; 4 \u2013 the disc sections are located in front of cultivator paws.",
        "texts": [
            " (61) It are calculated the periods of free vibrations of the mechanical system for four variants of the placement of operating bodies on the frame of the tillage unit with the purpose to determine the moving stability of the of the machine-tractor aggregate: 1 \u2013 flat-cutting paws are located in front of the disc sections; 2 \u2013 disc section are located in front of the flat-cutting paws; 3 \u2013 the cultivation paws are located in front of the disk sections; 4 \u2013 disc sections are located in front of cultivator paws. The damped nature of vibrations due to the viscosity of the soil did not take into the calculations. For all variants of the arrangement of the operating bodies of the mounted tillage unit the values of the free vibration period of this system are calculated by the formula \ud835\udf0f = 2\ud835\udf0b \ud835\udc58 . (62) The graphics of variation in the turning angle \ud835\udf11\ud835\udc42\ud835\udc43 (when the step of change \ud835\udf11\ud835\udc42\ud835\udc43 = 1\u00b0) of the tillage unit are shown in figure 2, when the machine-tractor aggregate is moving during tillage obtained as a result of calculations. We consider that the more shortly the time to the first crossing of the 0t-axis, the quicker the comeback of the operating bodies of the mounted tillage unit to the tolerance zone of the depth of tillage will occur and consequently the more steady moving of the unit. The time of the vibrations period for the first variant (the flat-cutting paws place in front the disk sections) was 1.213 s which is less than for the second variant (the disk sections locate in front the flat- ESDCA 2021 IOP Conf"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002047_978-1-4471-5110-4_6-Figure6.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002047_978-1-4471-5110-4_6-Figure6.13-1.png",
        "caption": "Fig. 6.13 Example 6.5",
        "texts": [
            " The partial derivative of the function V (x, y) with respect to x and y are \u2202V \u2202x = \u2212 (mg \u2212 k R) x\u221a R2 \u2212 x2 \u2212 y2 , \u2202V \u2202y = \u2212 (mg \u2212 k R) y\u221a R2 \u2212 x2 \u2212 y2 . The equilibrium positions of the particle are obtained from \u2202V \u2202x = 0, \u2202V \u2202y = 0. In MATLAB the equilibrium positions are obtained with: dVxydx = simple(diff(Vxy,x)); dVxydy = simple(diff(Vxy,y)); xe=solve(dVxydx,x); ye=solve(dVxydy,y); ze=solve(xe\u02c62+ye\u02c62+z\u02c62-R\u02c62,z); The results for the equilibrium positions are M1(0, 0, R) and M2(0, 0,\u2212R). Example 6.5 A particle P of mass m is on a circle of radius R as shown in the Fig. 6.13. The circle is on a vertical plane xy. Find the equilibrium positions of the particle. Solution The independent variable is the angle \u03b8. The position of the particle P is x = R*cos(theta); y = yN+R+R*sin(theta); r_ = [x y]; where yN is the y coordinate of the lower end N of the circle. The gravity is the only force acting on the particle and the potential energy is calculated with: dr_=diff(r_,theta); G_ = [0 -m*g]; V = -int(G_*dr_.\u2019); fprintf(\u2019V=%s + C\\n\u2019, char(V)) The MATLAB expression for the potential energy is: V=R*g*m*sin(theta) + C where C is a constant of integration. The equilibrium positions are calculated from the equation: dV = diff(V,theta); thetae=solve(dV,theta); theta1=thetae; theta2=theta1+pi; The equilibrium position are the points M and N as shown in Fig. 6.13: theta1 = pi/2 and theta2 = (3 \u2217 pi)/2. The equilibrium stability is verified with the second derivative of the potential energy: d2V = diff(dV,theta); d2V1=subs(d2V,theta,theta1); d2V2=subs(d2V,theta,theta2); and the MATLAB results are: d2V/d(theta)\u02c62=-R*g*m*sin(theta) for theta1 => d2V/d(theta)\u02c62=-R*g*m for theta2 => d2V/d(theta)\u02c62=R*g*m The equilibrium position \u03b8 = 3 \u03c0/2, position N , is a stable equilibrium because d2V/d\u03b82 = Rgm is positive. 6.6 Problems B A C 1 2 k \u03b8 m, l m, l F A B A C \u03b8 m, l l k k Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000982_9780471740360.ebs1320-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000982_9780471740360.ebs1320-Figure12-1.png",
        "caption": "Figure 12. Free-body diagram of a body segment. Sagittal plane view. D\u00bcdistal end of the segment. P\u00bcproximal end of the segment. Ry, Rx\u00bc reaction forces at the distal (D) and proximal (P) ends. Mz\u00bcmoment about the distal (D) and proximal (P) ends. Izpazp\u00bcmoment caused by resistance to segment inertia during rotation about the proximal end. maxcom, maycom\u00bc translational forces acting on the segment at the center of mass. mg\u00bcweight of the limb.",
        "texts": [
            ", the shoulder has over 16 muscles with larger muscles requiring more than one force vector) exist than knowns. In a link-segment model, each segment acts independently under the influence of reaction forces, muscle moments, inertial forces, and gravitational forces. Freebody diagrams are drawn for each segment and the net muscle moments and joint forces computed using 3-D force and moment balance equations: X Fx \u00bcMax X Fy \u00bcMay X Fz \u00bcMaz; \u00f034a\u00de X Mx \u00bc Ixax X My \u00bc Iyay X Mz \u00bc Izaz: \u00f034b\u00de A free-body diagram of a body segment is shown in Fig. 12. More complex biomechanical models have been developed and a few implemented in commercial software packages designed for clinical and rehabilitation settings. Some biomechanical measurement systems (e.g., motion analysis systems) have a built-in analysis feature that, after collecting the data, can solve the inverse dynamics solutions, display the graphs, and write data to a file. Computer software can be helpful for visualizing muscle length changes and real-time joint forces and moments as a result of the movement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002978_detc2011-48760-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002978_detc2011-48760-Figure7-1.png",
        "caption": "Figure 7. Tooth profile of gear modeled using 3D-CAD system.",
        "texts": [],
        "surrounding_texts": [
            "The numerical coordinates on the tooth surfaces Xg, Xg' Xp, and Xp' of the straight bevel gears were calculated based on the concept in the previous section. The determined coordinates are changed by the phase of one pitch after the tooth surfaces Xg, Xg' Xp, and Xp' are calculated. This process is repeated and produces the numerical coordinates on other right and left tooth surfaces. When the range of the existence of the workpiece that is composed of the root cone, face cone, heel, and toe etc. is indicated, the straight bevel gear is modeled [10]. Figures 7 shows the tooth profiles of the gear modeled using a 3D-CAD system based on the calculated numerical coordinates. The tool pass is calculated automatically after checking tool interferences, choosing a tool, and indicating cutting conditions. Therefore, the CAM process can be realized [11]. When the numerical coordinates of the tooth surfaces are calculated, the tooth surfaces are estimated by the smoothing of a sequence of points, removal of the profile of undercutting, offset of tool radius, and generation of NURBS surface. Moreover, the calculations of intersecting curved lines of right and left tooth surfaces, and sectional curved line, the approximation of straight line are conducted. The approach escape is added in order to avoid the interference. When the attitude of the tool and coordinate transformation are conducted, NC data and IGES (Initial Graphics Exchange Specification) data for the machining and display can be obtained. The pinion tooth profile is also modeled in the same manner as that of the gear."
        ]
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.69-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.69-1.png",
        "caption": "Fig. 2.69 4PRPaPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology P||R\\Pa\\\\Pa",
        "texts": [
            " 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No. 58 T ab le 2. 4 L im b to po lo gy an d co nn ec ti ng co nd it io ns of th e fu ll y- pa ra ll el so lu ti on s w it h no id le m ob il it ie s pr es en te d in F ig s. 2. 73 , 2. 74 , 2. 75 , 2. 76 , 2. 77 , 2. 78 ,2 .7 9, 2. 80 ,2 .8 1, 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure14-1.png",
        "caption": "Figure 14. Normalised Stress at RSFB region (Linear analysis)",
        "texts": [
            " Hence, FE model needs to prepared in such a way that no penetration should happen between the components Surface to surface contact is defined for the following contact pairs \u2022 RSFB bracket & External flitch \u2022 External Flitch & FSM \u2022 FSM & Internal flitch \u2022 Internal Flitch & RSFB Cross member Local non-linear analysis with contact has been carried out for the identified critical locations and compared the results between linear analysis and contact analysis. Figure 12 and 13 shows comparative displacement plot at RSFB region between linear and local non-linear analysis respectively. It is observed from Figure 12 that external flitch top flange is penetrating inside FSM top flange as contact is not defined in linear analysis. This is not so in the case of local non-linear analysis as contact is defined and are shown in Figure 13. Figure 14 and 15 shows comparative stress plot at RSFB region between linear and local non-linear analysis respectively. In linear analysis, high stress is observed in the fillet region of External flitch due to load transfer only through bolts and non-consideration of contact between External flitch top flange and FSM top flange. This is not so in case of local non-linear analysis as contact is defined. Figure 16 and 17 shows comparative stress plot at RSMB region between linear and local non-linear analysis respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001325_0-387-22644-3_31-Figure30.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001325_0-387-22644-3_31-Figure30.2-1.png",
        "caption": "FIGURE 30.2. Illustration of the coordinate system used to describe a spring-mass system. The mass is allowed to slide freely back and forth with no friction.",
        "texts": [
            " 1Robert Hooke (1635\u20131703) was a true general scientist. While holding a chair in geometry for most of his career, he also made numerous important scientific observations and worked as an architect and surveyor. We model a spring that has one end attached to a wall and the other end to a mass m that is allowed to slide freely back and forth on a table, neglecting friction. We choose coordinates so that when the spring is at its rest position the mass is located at s = 0 and s > 0 corresponds to stretching the spring to the right (see Fig. 30.2). In this coordinate system, Hooke\u2019s law for the force F reads, F = \u2212ks, (30.1) where the constant of proportionality k > 0 is called the spring constant. Combining Hooke\u2019s law with Newton\u2019s Law of Motion yields the equation m d2s dt2 = \u2212ks (30.2) determining the motion of the mass. This equation is usually rewritten as s\u2032\u2032 + \u03c92s = 0, \u03c9 = \u221a k m . (30.3) To describe specific solution, we also give some initial conditions at time t = 0. We can specify the initial position of the object, s(0) = s0, and the initial velocity, s\u2032(0) = s1, for example"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002009_978-3-319-01228-5_4-Figure4.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002009_978-3-319-01228-5_4-Figure4.10-1.png",
        "caption": "Fig. 4.10 Underactuated manipulator with passive joint",
        "texts": [
            "108) and the driven internal dynamics (4.107) for the unactuated states qu, q\u0307u reduce to the second order differential equation ( Muu( yd , qu) \u2212 MT au( yd , qu) ) q\u0308u = gu( yd , y\u0307d , qu, q\u0307u) \u2212 ku( yd , y\u0307d , qu, q\u0307u) \u2212 MT au( yd , qu) y\u0308d . Again, the inverse model consists of three parts which are shown schematically in Fig. 4.9. In the following, the choice of such a linearly combined output is demonstrated exemplarily for a manipulator with two active and one passive rotational joint, see Fig. 4.10. The length of the links are l1, l2, l3. The manipulator is described by the generalized coordinates q = [\u03b11,\u03b12,\u03b2]T , whereby \u03b2 denotes the unactuated coordinate. Input torques T1, T2 act in direction of the actuated coordinates \u03b11,\u03b12, respectively. Link 3 is connected by a passive joint to link 2 which is supported by a spring. The control goal is tracking of the desired trajectory re f d of the end-effector position re f = [ l1 sin(\u03b11) + l2 sin(\u03b11 + \u03b12) + l3 sin(\u03b11 + \u03b12 + \u03b2) \u2212l1 cos(\u03b11) \u2212 l2 cos(\u03b11 + \u03b12) \u2212 l3 cos(\u03b11 + \u03b12 + \u03b2) ] ",
            "109) Assuming a stiff spring, the \u03b2 angle remains small. Then, the linearly combined output y = [ y1 y2 ] = [ \u03b11 \u03b12 + \u0393 \u03b2 ] , (4.110) can be used to approximately describe the end-effector position re f so that re f \u2248 re f ap = [ l1 sin(\u03b11) + (l2 + l3) sin(\u03b11 + \u03b12 + \u0393 \u03b2) \u2212l1 cos(\u03b11) \u2212 (l2 + l3) cos(\u03b11 + \u03b12 + \u0393 \u03b2) ] = [ l1 sin(y1) + (l2 + l3) sin(y1 + y2) \u2212l1 cos(y1) \u2212 (l2 + l3) cos(y1 + y2) ] . (4.111) Thereby, the output y2 = \u03b12 + \u0393 \u03b2 can be geometrically interpreted as an auxiliary angle to approximate the end-effector point, see Fig. 4.10. Then, instead of tracking the end-effector position re f , the output y can be tracked. The desired trajectories yd of the output is computed from the desired trajectory re f d and (4.111), using one of the inverse kinematics techniques mentioned in Sect. 4.1.1. Due to the approximation of the end-effector position by the linearly combined output it is obvious that a small tracking error always occurs for \u03b2 = 0. In order to determine the value \u0393 , a Jacobian linearization of the end-effector position around a nominal trajectory qn = [\u03b11,n,\u03b12,n,\u03b2n]T of the generalized coordinates is performed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003002_amr.291-294.3282-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003002_amr.291-294.3282-Figure2-1.png",
        "caption": "Fig. 2. Finite element meshing at contact area",
        "texts": [
            " There are three types of elements were chosen to simulate rail-wheel contact problem, the element SOLID45, which is used for the 3-D modeling of solid structures and defined by eight nodes having three degrees of freedom at each node, for rail and wheel; the contact element CONTA174 for wheel contact area to represent contact and sliding between rail and wheel; and the target element TARGE170 for rail contact area as the \u2018target\u2019 surface for the associated contact elements, respectively. In order to improve the accuracy of contact calculation, all the elements in contact area have been subdivided. The contact elements for line-contact are shown in Fig. 2. The material properties for both wheel and rail are: E = 210 GPa, \u03bd = 0.3, \u03c1 = 7900 kg/m 3 , g = 9.8 m/s 2 , the friction coefficient is 0.3. The material of wheel is 42CrMo with surface hardening. An elasto-plastic stress-strain constitutive model is used for modeling the hardened surface material, where \u03c3s = 883 Mpa, Ep = 0.1E. In order to compare the numerical results, the elastic Hertz theoretical results are also calculated, three different wheels with different hardening depths are applied, and three loads has been applied vertically to all cases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002183_b978-0-7506-8496-5.00011-7-Figure11.8-1.png",
        "caption": "Figure 11.8: Fasteners with highest accelerations.",
        "texts": [
            " The accelerations versus frequency in three perpendicular directions at each fastener location under each of the above-mentioned excitations were computed. A typical response curve in a frequency range 0e50 Hz is illustrated in Figure 11.7. The maximum peak amplitude in each response curve was used as a performance metric of squeak and rattle implications. In essence, fasteners that experience high accelerations in a vehicle tend to get loose over time easily, which in turn causes squeak and rattle problems. The fastener locations with very high accelerations are shown in Figure 11.8 while details (e.g. critical excitations, maximum response peak, peak frequency) related to their responses are given in Table 11.3. The squeak and rattle concerns identified in vehicle tests are also given in this table. It can be found in this table that fasteners with excessively high acceleration values coincide with those exhibiting squeak and rattle concerns identified in vehicle tests. This provides the justification for using fastener accelerations as an indicator of IP squeak and rattle performance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure2-1.png",
        "caption": "Figure 2 \u02d8 Castor multibody model",
        "texts": [
            " In the present paper a new set of experimental tests are presented with the intent to identify the main vibration modes of the system. The tests are preceded by the description of the multibody model adopted to guide the experimental investigation. To predict the dynamic behaviour of the castor under test, a multibody model was developed with a commercial code (MSC VisualNastran). The geometry of the caster has been modelled by means of a CAD code (SolidWorks) to take into account its inertia characteristics and the constraints existing between the castor parts (Fig. 2). The model is constituted by the following rigid bodies: the headstock with the fork plate and the handlebar rigidly connected to the fork tubes; two sliders, rigidly connected to the wheel spindle and constrained to the fork tubes by means of a revolute joint (these bodies are also connected by means of a linear spring damper); the wheel is connected to the wheel spindle by means of a revolute joint. The headstock is constrained to the ground by means of a dummy body that allows the castor to translate; the vertical and the lateral translations are counteracted by the castor weight and by an horizontal spring, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001771_978-94-007-5313-6_6-Figure6.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001771_978-94-007-5313-6_6-Figure6.7-1.png",
        "caption": "Fig. 6.7 Surface Plasmon-coupled cone of emission for fluorophores near a metallic film (Reproduced from Lacowicz [1] by permission granted by Springer Science and Business Media)",
        "texts": [
            " The effect of monQm is larger for fluorophores with low quantum efficiency Q0, since m has no effect when Q0 D 1: The SPCE process is illustrated in Fig. 6.6 [1], in which F represents a fluorophore set above a continuous silver film 50 nm thick. The fluorescence emission from the fluorophore is not reflected, but passes through the film. The spatial pattern of the fluorophore emission is nearly isotropic, but the emission seen through the film is at an angle \u2122F with respect to the normal. The sample being symmetrical about the normal z-axis, the emission through the film occupies a cone around the z-axis (see Fig. 6.7). The emission on the cone, about a half of the total emission (the rest being free space emission away from the film) is called Surface-Plasmon Coupled Emission (SPCE). This name reflects the fact that the emission in the cone is radiated by the plasmons. The following can be said: 1. The spectral distribution of the cone emission is the same as that of the fluorophore. 2. The plasmons are not the result of RET because the distance at which SPCE takes place (200\u20132,000 \u00c5) are too large for RET, for which the F\u00f6rster distance is 50 \u00c5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002278_2013-01-1491-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002278_2013-01-1491-Figure2-1.png",
        "caption": "Figure 2. 6 speed manual Transmission for medium duty Commercial Vehicle",
        "texts": [
            " In this paper we will discuss on detail about the optimization of radial gap between gear bore diameter and outside diameter of the shaft through gear tilt stack and also discuss about the runout variation because of the tilt in the gear due to the radial gap. A 6 speed manual transmission of medium duty commercial vehicle has been taken and mathematically analyzed to check the possible gear tilt and run out due to radial clearance between the gears and needle roller bearing. A transmission model has been constructed on the Romax designer software and compares the gear strength variation due tilting of the gear. A 6 speed manual transmission shown in Figure 2, has been taken to calculate maximum radial gap, maximum tilting of gear and maximum runout for 2nd gear pair, through tolerance stack up analysis. The detailed description and the method to perform tolerance stack up is described below. 6 speed manual transmission of the medium duty commercial vehicle has been taken to analyze the gear tilting angle in milli radians through tolerance stack up analysis. The method followed for the tolerance stack up analysis is described in various steps. Before to start the tolerance study, list out the dimensions to be used in stack path and also draw the sketches to explain the path and gear tilt"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002780_s10851-011-0286-y-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002780_s10851-011-0286-y-Figure9-1.png",
        "caption": "Fig. 9 Aibo head\u2019s degrees of freedom. Lateral and front views",
        "texts": [
            "2 Transformation bIg Between Ig and Ib The first component of the transformation bIg is the rotation matrix R0, built from the three director vectors defining the axes of Ig : R0 = \u239b \u239dg2 \u2212 g1 g3 \u2212 g1 (g3 \u2212 g1) \u00d7 (g2 \u2212 g1) 0 0 0 0 1 \u239e \u23a0 , (14) and the second is the translation from the origin of Ib to the origin of Ig, T0 = ( I3\u00d73 g1 0 1 ) . (15) So, composing the two transformations we finally obtain the matrix transformation from Ig to Ib , bIg = T0R0. (16) 4.2.3 Transformation cIb Between Ib and Ic Figure 9 shows the neck and head geometric parameters. There are two tilting degrees of freedom of the Aibo, denoted \u03b8tilt and \u03b8nod . The first corresponds to the neck base pivoting over the chest, the second allows corresponds to the head moving vertically at the head-neck articulation. The third degree of freedom, called \u03b8pan, allows a rotation perpendicular to the tilt, moving the head from side to side. The transformation between the camera and body reference systems is the following composition: bIc = T3R2T2R1T1, (17) where T1 is the translation from the camera base to the top of the neck, R1 accounts for the nod and pan head rotations, T2 is the translation from the neck top joint to the neck base joint, R2 accounts for the neck-head tilt rotation, and T3 is the translation from the neck base to the body mass center"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003778_978-3-642-17234-2_4-Figure4.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003778_978-3-642-17234-2_4-Figure4.16-1.png",
        "caption": "Fig. 4.16 Geometrical representation of nutation for various conditions.",
        "texts": [
            "18)1 in the form _u \u00bc b au 1 u2 ; we realize that _u does not change its sign, meaning that the axis of the top performs a precession motion about the vertical line passing through O. If _u \u00bc 0; the precession is monotonic. When u reaches the values u1 or u2, in view of (4.6.28), tan a = ?, i.e. the curve u \u00bc u\u00f0h\u00de on the sphere is tangent to the parallel circles h \u00bc h1; h \u00bc h2: The axis of the top performs a periodic motion of Fig. 4.15 Parameters describing the case (b) f\u00f0u0\u00de \u00bc 0. lifting and descending. This motion is done on the background of the precession motion and it is called nutation (Fig. 4.16a). In the second case, the curve u \u00bc u\u00f0h\u00de is also tangent to the parallel circles h \u00bc h1; h \u00bc h2; but _u changes its sign for u \u00bc u0; meaning that the directions of the precession on the circles h \u00bc h1; h \u00bc h2 are opposite (Fig. 4.16b). Let us, finally, analyze the third case. Putting b \u00bc au0 in (4.6.27), we have: f \u00f0u\u00de \u00bc \u00f01 u2\u00de\u00f0a bu\u00de a2\u00f0u0 u\u00de2 \u00f04:6:29\u00de and, since either u0 \u00bc u1; or u0 \u00bc u2; it follows that f \u00f0u0\u00de \u00bc \u00f01 u02\u00de\u00f0a bu0\u00de \u00bc 0: This leads to a \u00bc bu0, and (4.6.29) takes the form: f \u00f0u\u00de \u00bc \u00f0u0 u\u00de\u00bdb\u00f01 u2\u00de a2\u00f0u0 u\u00de : \u00f04:6:30\u00de Writing (4.6.30) for the two roots u1 and u2, we have: f\u00f0u1\u00de \u00bc \u00f0u0 u1\u00de\u00bdb\u00f01 u2 1\u00de a2\u00f0u0 u1\u00de \u00bc 0; f\u00f0u2\u00de \u00bc \u00f0u0 u2\u00de\u00bdb\u00f01 u2 2\u00de a2\u00f0u0 u2\u00de \u00bc 0: As we can see, if in one of these equations the parenthesis vanishes, in the other the bracket becomes zero",
            " Since 1 - u2 [ 0, it results that either u0[ u1; or u0[ u2: But u0[ u2 does not have any sense, so that the only remaining possibility is u0[ u1: This inequality is satisfied only by u0 \u00bc u2; i.e. u0 may coincide only with the biggest of the two roots u1; u2: At the same time, (4.6.28) leads to tan a = 0, for u \u00bc u0 \u00bc u2; which means that the angle a, made by the curve u \u00bc u\u00f0h\u00de at the intersection points with the parallel circle h = h2, is zero. Since _u does not change its sign in the interval \u00f0u1; u2\u00de; we conclude that these points are turning points (Fig. 4.16c). In order to explain the motion of nutation, let us take as initial conditions: h\u00f00\u00de \u00bc h0 6\u00bc 0; p; _h\u00f00\u00de \u00bc 0; _u\u00f00\u00de \u00bc 0: \u00f04:6:31\u00de The relations (4.6.16), (4.6.19) and (4.6.20) then lead to \u00f0E1\u00det\u00bc0 \u00bc E 1 2 I03x 0 3 2 t\u00bc0 \u00bc Mgl cos h0 \u00bc const. \u00f04:6:32\u00de In the light of this result, let us now give a physical interpretation of the equation of conservation (4.6.17). It shows that, if at the initial moment t = 0, the quantity E1 satisfies Eq. (4.6.32), then the potential energy must diminish with the growth of _u; _h: The angle h0 is then precisely the minimum value h2 of h, i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001650_978-1-4419-8113-4_16-Figure16.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001650_978-1-4419-8113-4_16-Figure16.3-1.png",
        "caption": "Fig. 16.3 Bounce of a ball incident obliquely on a horizontal surface. 1 is the angle of incidence and 2 is the angle of reflection. N is the vertical force on the ball and F is the horizontal friction force. v1 is the incident speed and v2 is the bounce speed",
        "texts": [
            " A tennis player therefore needs to be careful, when tilting the racquet, to tilt it in the right direction. A baseball player tends to get whatever comes, unless he is skillful enough to strike the ball exactly where he wants to. We will return to this later when we examine whether a curveball (incident with topspin) can be struck farther than a fastball (incident with backspin). When a ball without spin or with topspin is incident at an oblique angle on a flat, heavy, horizontal surface, the ball will bounce with topspin, as indicated in Fig. 16.3. Provided the surface is much heavier and stiffer than the ball then motion of the surface itself can be ignored and the bounce is determined mainly by the properties of the ball. Nevertheless, there is one property of the surface that does influence the bounce, and that is the smoothness or roughness of the surface. If the surface is slippery then the friction force on the ball will be relatively small. If the surface is rough then there will be a large friction force on the ball. The friction force on the ball acts in a direction parallel to the surface and has two effects",
            " The only way to counter the hype is to take careful measurements of ball speed and spin to determine whether there is any substance to the manufacturer\u2019s claims. Some progress has been made in this direction but a lot more still needs to be done. Suppose that a ball is incident obliquely on a horizontal surface, at speed v1, and bounces at speed v2, as shown in Fig. 16.5. The horizontal components of the ball speed before and after the bounce are vx1 and vx2, respectively. The latter speeds refer to the speed of the ball center of mass (CM). Suppose also that the ball is incident with topspin and bounces with topspin, as shown in Fig. 16.3, with angular speeds !1 and !2, respectively. A point at the bottom of the ball will have a lower horizontal speed than the CM since the bottom of the ball is rotating backward. The horizontal speeds at the bottom of the ball are s1 D vx1 R!1 before the bounce and s2 D vx2 R!2 after the bounce, where R is the radius of the ball. If the contact point has a vertical speed vy1 before the bounce, and vy2 after the bounce then we define ey D vy2=vy1 as the COR in a direction perpendicular to the surface",
            " I have not investigated this effect in enough detail to figure out why this happens. It seems that the ball surface is more elastic when the ball is relatively new and that repeated bounces act to harden the surface and to reduce its elasticity in a direction parallel to the surface. The end result is that a baseball can bounce with topspin when it is incident with backspin, but only for the first few bounces. After ten or more bounces the ball tends to bounce without any spin at all when it is incident with backspin. Appendix 16.1 Ball Bounce Calculations (a) Sliding In Fig. 16.3, we show a ball incident at speed v1 and bouncing at speed v2 off a heavy surface. The horizontal component of v1 is vx1 D v1sin 1 and the vertical component is vy1 D v1cos 1. Likewise, the components of v2 are vx2 D v2sin 2 and vy2 D v2cos 2. When the ball is sliding along the surface, a friction force F D N acts on the bottom of the ball in a direction parallel to the surface. N is the normal reaction force acting on the ball in a direction perpendicular to the surface, and is the coefficient of sliding friction (COF)",
            " For a baseball on wood or aluminum, is typically about 0.4 or 0.5. If F and N are taken as average forces during the bounce, and if the ball remains in contact with the surface for a time T then the change in the horizontal and vertical components of the ball speed during the bounce are given by F D m dvx dt D m.vx1 vx2/ T (16.1) and N D m dvy dt D m.vy1 C vy2/ T ; (16.2) where m is the ball mass. To avoid complicating the issue with negative numbers and signs, we have assumed here that all quantities are positive in the direction shown in Fig. 16.3. In particular, the ball reverses direction in the y direction, but we can still take vy1 to be a positive number if we want to. If the ball is incident in the vertical direction at speed vy1 D 5 m s 1 and bounces at speed vy2 D 2 m s 1, then the change in speed is 7 m s 1, not 3 m s 1, since the ball reverses direction. If the ball bounced with vy2 D 5 m s 1, then the change in vertical speed would be 10 m s 1, not zero. The coefficient of restitution, ey , for a bounce on a heavy surface, is defined by ey D vy2 vy1 (16",
            " For a bat and ball collision, the angle of incidence is increased when the batter strikes the ball near the bottom of the ball rather than striking the middle of the ball. One way to increase ex would be to coat the bat with a soft, flexible material like rubber. In that case, extra elastic energy would be stored in the rubber and then given back to the ball, increasing the values of both ey and ex . The ball would not only bounce at higher speed but it would also spin faster. That is why table tennis bats have a rubber surface. (c) Oblique Bounce Off a Light Surface Suppose that the surface in Fig. 16.3 is only slightly heavier than the incident ball. Then the ball will bounce off the surface at reduced speed since the ball transfers some of its kinetic energy to the surface. That is essentially the situation encountered when a ball strikes a stationary bat or a tennis racquet. If the ball makes a head-on collision with the bat then conservation of momentum for the collision, plus an estimate of the COR, tells us the recoil speed of the bat. We can then proceed as in Chap. 9 to calculate the bounce speed of the ball and the recoil speed of the bat. However, if the ball does not strike the bat head-on then the ball will be deflected sideways by the curved surface of the bat, in which case we can treat the collision as an oblique bounce, as shown in Fig. 16.3. A similar situation occurs in tennis when the ball strikes the strings at an oblique angle and the racquet head recoils. Recoil motion of the bat or the racquet adds to the complexity of the problem, but there is a simple way around the problem. That is, we ignore the recoil motion and just measure what happens to the ball. We saw in Chap. 9 that it is very useful to define an apparent coefficient of restitution, eA, or bounce factor, q describing the ratio of the ball speed after the collision to that before the collision"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002503_isse.2011.6053543-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002503_isse.2011.6053543-Figure2-1.png",
        "caption": "Fig. 2. Model of substrate with thin layer of DPP.",
        "texts": [
            " Suitable polymerizable groups are halogen, hydroxyl, trifluoromethylsulfonate, or aldehyde groups [1,4]. 978-1\u20134577\u20132112\u20130/2011/$ 26.00 \u00a9 2011 IEEE 22 34th Int. Spring Seminar on Electronics Technology In our case, the molecules of DPP exist as single compact molecular units, they are not bound to each other in a polymer chain. 3.1. Substrate and electrode system The substrate used for experiments contains up to 99 % of alumina. Golden interdigital electrodes are created by \u201clift off\u201d method on the top of ceramic substrate (Fig. 2). The electrodes fingers are 27 \u03bcm wide and the width of the insulation gap is 25 \u03bcm. Thickness of the electrodes is 400 nm. The material of electrodes ensures that organic material will have any corrosive impact on them. This interdigital electrode system is suitable for simple creating organic layers on the top of ceramic substrate by printing technique. Printing technology seems to be a leading technique for cheap large scale produce of such devices. 3.2. Organic materials Four types of DPP were chosen for this experiment (Fig",
            " The relative humidity was changed from 30 % to 85 % simultaneously with temperature from 30 \u00b0C to 85 \u00b0C (Fig. 5). These rapid changes were repeated twelve times. This test cycle was used in order to get the long-term stability of electrical parameters. The impedance of these layers were measured by using four wires method by means of precision LCR meter at frequency of 1 kHz and voltage of 1 V like in previous test. Impedance-humidity characteristic of DPP\u2019s layers at the temperature of 30 \u00b0C are shown in the figure 2. 978-1\u20134577\u20132112\u20130/2011/$ 26.00 \u00a9 2011 IEEE 24 34th Int. Spring Seminar on Electronics Technology For this test was used standard climatic cycle (Fig. 4). All DPPs show the dependence of their electrical properties on the level of relative humidity. DPP2 shows a suitable curve path. The values of the impedance stay within the range of three orders of magnitude. Additionally, the maximum value of impedance is lower than 1 MOhm. This is suitable for further measurement and evaluation. The next tests have been done only for DPP2, showing the best results"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003464_amm.397-400.176-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003464_amm.397-400.176-Figure2-1.png",
        "caption": "Fig. 2 Planet Carrier Design",
        "texts": [
            " These could be obtained at the correct size for a reasonable price to allow for a number of different faults to be seeded. Deep groove ball bearings were also used for supporting the planet carrier on the driven-shaft. The calculated load on the planet bearings, at a torque load of 30 Nm, is only 234 N in ideal conditions. Under uneven load distribution the force could be as high as 460 N. This is well below the manufactures rated load of 1320N. The most critical part of the gearbox design and manufacture was the planet carrier shown in Fig. 2. This single part interfaces with almost every other component in the system and transmits torque from the first stage of the gearbox to the second. Another important function of the planet carrier was to align the planet gears axially and radially with the driven shaft while rotating independently of this shaft. The planet carrier transmitted torque from the spur gear to the planet gears and each planet gear was cantilevered off bearing pins. It was therefore essential that the carrier assembly was not only strong enough to support the applied loads but also stiff enough to allow only minimal deflections under load"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002692_acc.2012.6315666-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002692_acc.2012.6315666-Figure1-1.png",
        "caption": "Fig. 1: Configuration described in Lemma 1 for the case when |\u03b1i + \u03b3i|> \u03c0 2 .",
        "texts": [
            " Then, (7) can be written as ri =\u2212 \u2211 j\u2208Ni(G) \u03b2i j(qi \u2212q j) (9) It follows from (3) and connectivity preservation that \u03b2i j > 0. Moreover, since qi is the farthest point from P, the circle centered at P with the radius \u2225P\u2212 qi\u2225 contains all agents, and hence (P\u2212qi) T (q j \u2212qi)\u2265 0, for all j \u2208 Nn. Therefore, (P\u2212qi) T ri = \u2211 j\u2208Ni(G) \u03b2i j(P\u2212qi) T (q j \u2212qi) \u2265 0 (10) which in turn implies that |\u03b3i| \u2264 \u03c0 2 . On the other hand, d dt \u2225P\u2212qi(t)\u2225 = \u2212 (P\u2212qi) T \u2225P\u2212qi\u2225 q\u0307i = \u2212 (P\u2212qi) T \u2225P\u2212qi\u2225 Rot(\u03b1i)ri cos\u03b1i = \u2212\u2225ri\u2225cos(\u03b1i + \u03b3i)cos\u03b1i = \u2225ri\u2225sin(|\u03b1i + \u03b3i|\u2212 \u03c0 2 )cos\u03b1i (11) This is illustrated in Fig. 1 for the case where |\u03b1i + \u03b3i|> \u03c0 2 . It follows from |\u03b1i| \u2264 \u03c0 2 and |\u03b3i| \u2264 \u03c0 2 that \u2212\u03c0 2 \u2264 |\u03b1i + \u03b3i|\u2212 \u03c0 2 \u2264 \u03c0 2 . If \u2212\u03c0 2 \u2264 |\u03b1i + \u03b3i|\u2212 \u03c0 2 \u2264 0, then it is concluded from (11) that d dt \u2225P \u2212 qi(t)\u2225 \u2264 0 and the proof is complete. If 0 < |\u03b1i + \u03b3i|\u2212 \u03c0 2 \u2264 \u03c0 2 , then on noting that sinx < x for any x \u2208 (0, \u03c0 2 ], (11) yields d dt \u2225P\u2212qi(t)\u2225 \u2264 \u2225ri\u2225(|\u03b1i + \u03b3i|\u2212 \u03c0 2 )cos\u03b1i \u2264 rM(|\u03b1i + \u03b3i|\u2212 \u03c0 2 ) \u2264 rM(|\u03b1i|+ |\u03b3i|\u2212 \u03c0 2 ) \u2264 rM|\u03b1i| (12) which completes the proof. Lemma 2: Let \u03b1M = maxi\u2208Nn |\u03b1i(0)|, and TM = max{ln 2\u03b1M \u03c0 ,0}"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.23-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.23-1.png",
        "caption": "Figure 3.23 The aircraft control surfaces, the elevator, the ailerons, and the rudder",
        "texts": [
            "144) are very useful in the derivation of atmospheric flight dynamics plant as they can employ the closed-form results of small perturbation aerodynamic theories. The aerodynamic force and torque vectors can be separated into forces and moments arising due to relative velocity (both linear, v, and angular, \u03c9) of the rigid aircraft with respect to the atmosphere (equation (3.144)), and the control forces and moments generated by the control surfaces, namely the elevator, the ailerons, and the rudder, as shown in Figure 3.23. The control surfaces are generally deflected by small angles, thus the aerodynamic control forces and moments obey linear relationships with control surface deflections. Furthermore, the control surfaces are driven by actuators with a dynamic response about ten times faster than that of the vehicle\u2019s rigid body dynamics. Therefore, flow unsteadiness due to control surface deflections can often be neglected when considering rigid body aerodynamics. Hence, a quasi-steady approximation of aerodynamic loads linearly varying with control surface deflection is usually employed",
            " 6 When the vehicle\u2019s structural flexibility is taken into account, it is necessary to include the unsteady aerodynamic effects of control surface deflections, that is, dependence of aerodynamic forces and moments on deflection rates, \u03b4\u0307E, \u03b4\u0307A, \u03b4\u0307R. edge of a larger stabilizing surface in the (xy) plane, as shown in Figure 3.21. Elevator deflection, \u03b4E, creates a forward force, X\u03b4\u03b4E, downforce, Z\u03b4\u03b4E, and a control pitching moment, M\u03b4\u03b4E. A pair of control surfaces in the (xy) plane, located symmetrically about the axis (ox) and deflected in mutually opposite directions by angles \u03b41 and \u03b42 (Figure 3.23) are called ailerons that are used as roll control devices. The aileron deflection is the average of the two separate deflections, \u03b4A = (\u03b41 + \u03b42)/2, and is designed such that a control rolling moment, LA\u03b4A, is produced along with a much smaller, undesirable yawing moment, NA\u03b4A. A control surface mounted on a fin in the (xz) plane behind the (oy) axis is called the rudder. A rudder deflection, \u03b4R, creates a sideforce, YR\u03b4R, control yawing moment, NR\u03b4R, and a much smaller rolling moment, LR\u03b4R. The rudder is primarily used to correct a lateral flight asymmetry",
            " Finally, the derivatives Yp, Lr, Np are called cross-coupling derivatives and represent the aerodynamic coupling between roll, yaw, and sideslipping motions. They are primarily due to the wing and the vertical tail. In addition to the perturbation variables, we have the control inputs applied by the pilot/automatic controller in the form of aileron angle, \u03b4A, and rudder angle, \u03b4R. A positive aileron angle is one that creates a positive rolling moment, which is an upward deflection of the control surface on the right wing, and a downward deflection on the left, as shown in Figure 3.23. Similarly, a positive rudder deflection is one that produces a positive yawing moment, that is, a deflection to the right (Figure 3.23). By substituting equation (3.182) into equation (3.181) and adding the control force and moments, we have the following state-space representation for lateral-directional dynamics: \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03c3\u0307 \u03c8\u0307 \u03b2\u0307 p\u0307 r\u0307 \u23ab\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23ad = ALD \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03c3 \u03c8 \u03b2 p r \u23ab\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23ad + BLD { \u03b4A \u03b4R } , (3.183) where ALD = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0 0 1 tan \u03b8e 0 0 0 0 sec \u03b8e g cos \u03b8e U 0 Y\u03b2 mU Yp mU Yr mU \u2212 1 0 0 AL\u03b2 + BN\u03b2 ALp + BNp ALr + BNr 0 0 CN\u03b2 + BL\u03b2 CNp + BLp CNr + BLr \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 , (3.184) BLD = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0 0 0 YA mU YR mU ALA + BNA ALR + BNR CNA + BLA CNR + BLR \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 , (3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001944_9781118354162.ch13-Figure13.29-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001944_9781118354162.ch13-Figure13.29-1.png",
        "caption": "Figure 13.29 A three-electrode screen-printed platform for amperometric sensors. (a) Gold working electrode; (b) Ag/AgCl reference electrode; (c) gold auxiliary electrode; (d) silver output contacts. Reproduced with permission from [54]. Copyright 2008 Faculty of Health and Social Studies, University of South Bohemia, Czech R.",
        "texts": [
            " An overview of these fabrication methods is presented in Section 5.13). Briefly, screen printing is based on the application of a paste on a flat support through openings in a screen. Adhesion is imparted by subsequent thermal treatment. Conducting layers for electrodes are obtained using pastes including conducting microparticles of carbon or metal precursors (e.g., a metal oxide). During the thermal treatment, the metal oxide gives, by reduction, a metal film. A typical sensor platform produced by screen printing is shown in Figure 13.29. On a small size strip, the working, reference and auxiliary electrodes can be formed in a concentric configuration. Reference Ag/AgCl electrodes can be obtained by the anodic reaction of silver in HCl or by the chemical oxidation of silver in a FeCl3/KCrO3Cl solution. Connection to the output contacts is made by insulated silver tracks. Being inexpensive, screen-printed electrodes are suitable for fabricating disposable sensors. Owing to the versatility of this technique, screen-printed devices are widely used as platforms in electrochemical enzymatic sensors [52,53]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002153_cbo9780511760723.019-Figure17.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002153_cbo9780511760723.019-Figure17.5-1.png",
        "caption": "Fig. 17.5 An interdigitated electrode array.",
        "texts": [
            " As previously discussed, changing the shape and orientation of electrodes, in addition to modulating the frequency and phase applied, can give rise to dielectrophoretic particle trapping, dielectrophoretic sorting, electrorotation, and twDEP effects. We now explore specific geometries and their key experimental parameters. Interdigitated Electrode Array. The interdigitated electrode array is a common electrode configuration used in DEP studies. The electrode array consists of two sets of electrodes, grounded and energized, that alternate spatially. This creates a nonuniform field in the region of the electrode array that traps particles against a flow (Fig. 17.5). Castellated Electrodes. Castellated electrodes are similar to interdigitated electrodes. Rather than straight electrodes, however, the castellated electrode array consists of square-wave-shaped electrodes. These patterns are usually placed parallel to each other and create alternating regions of low and high electric field. Symmetric and offset configurations of castellated electrodes are shown in Fig. 17.6. Traps. Electrode-based DEP traps generally consist of geometries that trap single particles"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001817_9781118443293.ch7-Figure7.19-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001817_9781118443293.ch7-Figure7.19-1.png",
        "caption": "Figure 7.19 A resistive Wheatstone bridge.",
        "texts": [
            " This could be realised using the inverting amplifier shown in Figure 7.10, where the input voltage is fixed at Vref and the conductometric sensor (RS) replaces the input resistance Rin. The output of this circuit will be: Vout \u00bc Vref Rf RS : \u00f07:17\u00de Therefore, the output voltage is inversely proportionate to the resistance of the sensor, that is it is proportional to the conductance. Another common method for resistance to voltage transduction is the Wheatstone bridge and it is worth reviewing the principle behind this circuit. Referring to the circuit shown in Figure 7.19, if we assume the output potential Vo \u00bc 0 V then the voltage drops across R1 and R3 must be equal: VAD \u00bc VAB ) I1R1 \u00bc I2R3: \u00f07:18\u00de Similarly, the voltage drops across R2 and R4 must also be equal: VDC \u00bc VBC ) I1R2 \u00bc I2R4: \u00f07:19\u00de If we divide these together then we get an equation stating that the ratios of the resistors in one side of the bridge must equal those in the other half: R1 R2 \u00bc R3 R4 : \u00f07:20\u00de This is a balanced bridge circuit but a small difference in one resistor will unbalance the bridge so that Vo 6\u00bc 0V "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001996_1.5062329-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001996_1.5062329-Figure1-1.png",
        "caption": "Fig. 1 Different cuts and their associated machining parameters used to design diffident laser milling strategies, in particular (1) border/profile cuts; (2) hatching cuts; (3) laser spot diameter; (4) border cut track displacement (step-over); (5) hatch track displacement (step-over); (6) distance between the end of the hatch line and the innermost border cut.",
        "texts": [
            " The best machining results are achieved when a pulsed laser beam is strongly absorbed near the surface of the substraue material. The pulse duration affects the laser power delivered on the substrate surface and the physical mechanism of the ablation process. The laser wavelength combined with the pulse energy and duration are important factors that predefine the nature of the material-beam interactions and the final machining results. A range of machining strategies can be employed to remove sequentially layers of material and thus to produce complex 3D structures from 2D data obtained by slicing a 3D CAD model. Figure 1 shows the two different laser cuts and associated with them machining parameters that are used to design the laser milling strategies. For example, four different laser milling strategies that are commonly used to produce micro tooling cavities are shown in Figure 2 [2], in particular: (a) Random \u201chatching\u201d inside together with cuts along the border as shown in Figure 2(a); (b) Profile cuts only, Figure 2(b); (c) Hatching only, Figure 2(c); (d) The cuts have the same configuration as those for strategy (a), and the only difference is the tilted beam in the profile cuts in order to produce vertical walls, Figure 2(d)",
            " By applying etching technologies, especially by etching silicon wafers, very precise hollow microneedles can be produced; however, these manufacturing routes are often slow, costly and a limited choice of materials is available. In addition, silicon is not suitable for producing devices that are in direct contact with human tissues, while silicon oxide is brittle and there is a high risk of microneedles breaking off during their application. For this case study, the use of several manufacturing technologies for producing masters for replicating arrays of micro-needles (see Figure 1) was considered, including: micro-milling, laser milling (or laser ablation) using nano (ns) and pico-second (ps) lasers, Electro Discharge Machining (EDM) and electroplating. Figure 4 depicts the possible use of these technologies, especially how they can be combined for producing such masters. Laser milling, and especially ultra short pulse laser ablation, like ps laser machining, was considered as a preferred technology for manufacturing the microneedles mould insert due to its capability to produce high resolution features/structures and its potential to fabricate the necessary cavities for replicating needles with sharp tips in comparison to the other micro-manufacturing processes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003785_icnmm2013-73187-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003785_icnmm2013-73187-Figure8-1.png",
        "caption": "Fig. 8(a) Schematic of experimental system components",
        "texts": [
            " Figure 7 shows the magnetic flux density of the permanent magnet ring in Z direction. The maximum magnetic flux density was about 0.13(T) at a distance of 0.9 mm. In this figure, the magnetic flux density increases with the distance until a peak magnetic flux density is reached, and then the magnetic flux density decreases monotonically. A microtube was placed inside the hole of the magnet ring, i.e., on the Z axis and the motion of the ferrofluid in the Z direction was observed in the following experiments. Figure 8(a) shows a schematic diagram of the ferrofluid experiment. We made an experimental device consisting of a CCD camera, micrometer, permanent magnet ring, translation stage, LED light, personal computer, ferrofluid, and microtube. Figure 8(b) shows the detailed locations of the microtube and magnet ring. The microtube containing a ferrofluid was placed in the direction parallel to the magnet\u2019s central axis. The 3 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/26/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use ferrofluid was pulled by the magnetic force of the magnet. The position of the magnet ring was controlled by a micrometer. The magnet ring was pushed by the head of the micrometer until the ferrofluid slug in the microtube began to move"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002681_ijcnn.2012.6252849-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002681_ijcnn.2012.6252849-Figure2-1.png",
        "caption": "Figure 2. The Robot Partner : MOBiMac",
        "texts": [
            " If our system occurs a blind spot by dynamic objects in living space, the system compensates for missing information by using the robot partner with inner sensor. Therefore, the environment surrounding people and robots should have a structured platform for gathering, storing, transforming, and providing information. Such an environment is called informationally structured space (Fig.1) [12]. The most important task in the informationally structured space is the estimation of human position, state, and behavior for natural communication. The robot partner can use the estimated information at any time. We developed a robot partner; MOBiMac shown in Fig. 2. Two CPUs are used for the interaction with human and the control of the robotic behaviors. The robot has two servomotors, eight ultrasonic sensors, a laser range finder (LRF) and a CCD camera. An ultrasonic sensor can measure the distance to objects. Through these specifications, the robot partner should understand the meanings of the expression (actions) of the human. In our study, robot uses a LRF sensor in order to acquire external information (Fig. 3). The LRF can measure distances up to approximately 4,095 mm in 682 different directions the covering measurement range is 240\u00b0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001230_0470036427.ch3-Figure3.14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001230_0470036427.ch3-Figure3.14-1.png",
        "caption": "Figure 3.14 Voltage and lagging current phasor.",
        "texts": [
            " What we really care about are two aspects of the oscillating quantity we are representing: its magnitude, and its relative timing or phase in relation to another sinusoidal function, as, for example, in the phase difference between voltage and current. The magnitude is represented in the phasor diagram simply as the length of the arrow. But the phase shift is now also straightforward to represent, namely, as an angle within the circle in the complex plane. Consider the voltage and current waves in Figure 3.14 below. The current lags the voltage by 30 degrees, meaning that the peak current always occurs with a delay of one-twelfth of a cycle after the voltage peak. The notion that the voltage is zero at time zero and peaks at exactly 90 degrees was an arbitrary choice, amounting to when we decide to start counting time, and choosing that starting point so as to be convenient for computation. But once we decided on how to label the timing for voltage, we are committed to describing the timing of current properly in relation to that voltage timing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.61-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.61-1.png",
        "caption": "Fig. 2.61 4PaRPaP-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\R\\Pa\\kR",
        "texts": [
            "22n) The axes of the revolute joints connecting links 7 and 8 of the four limbs are parallel 42. 4PaPaPR (Fig. 2.54) Pa||Pa||P\\R (Fig. 2.22o) Idem No. 4 43. 4PRPaPa (Fig. 2.55) P\\R\\Pa||Pa (Fig. 2.22p) Idem No. 5 44. 4PPaRPa (Fig. 2.56) P||Pa\\R\\Pa (Fig. 2.22q) Idem No. 33 45. 4PPaPaR (Fig. 2.57) P||Pa||Pa\\R (Fig. 2.22r) Idem No. 4 46. 4PaPPaR (Fig. 2.58) Pa||P||Pa\\R (Fig. 2.22s) Idem No. 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002387_20130918-4-jp-3022.00053-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002387_20130918-4-jp-3022.00053-Figure1-1.png",
        "caption": "Fig. 1. Schematic model of the hovercraft.",
        "texts": [
            "00053 The system showed in this work is composed by a dynamical model of a hovercraft with a low-level tracking controller. The Hovercraft is a radio controlled vehicle that has three degrees of freedom, two associated with the movement in the plane of its centre of masses (x, y), and one more associated with the yaw angle . The forces of the thrusters are u1 and u2, l is the distance between the centre of the fan and the symmetry axis, V is the Velocity of the hovercraft that can be decomposed in surge u and sway v, and \u03a8 is the drift angle. (See fig. 1). Under simplifying symmetry assumptions the dynamic equations can be expressed in inertial frame as follows (Cremean, et al., 2002). cos( ) sin( ) x y x u x y u y r x v y v v F d v v F d v d (1) Were F=(u1+u2)/m and =(u1-u2)/m are normalized force and torque induced by two on/off thrusters that give asymmetric trust in both directions u1,2 -umin,0,umax , du and dv are hydrodynamic damping coefficients, and m, and J are the mass and moment of inertia of the hovercraft (plus added masses). These parameters have been experimentally obtained by measurements in a real laboratory RC Hovercraft system (Aranda et al"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.4-1.png",
        "caption": "Figure 4.4 Design of two-part cast housing of one-step reducer with cylindrical gears",
        "texts": [
            " In reducers with herringbone gears or bifurcation of power, one of the shafts (preferably the pinion shaft as the lighter element) should be secured on roller bearings. Bearings of this type permit axial play; as a result, during operation of the gear the pinion can self-aline relative to the wheel gear (under the action of axial forces on each side of the herringbone) and the load will be uniformly spread between the herringbone. An example of the correct design of a one-step reducer with cylindrical gears is shown in Figure 4.4 and the corresponding experiential dimensions in Table 4.6. T a b le 4 .6 A p p ro x im at e d im en si o n s o f ca st h o u si n g p u rs u an t to F ig u re 4 .4 a d d 1 C D h h 1 d 1 S ta n d ar d ce n tr e d is ta n ce 0 .0 3 a \u00fe5 m m (0 .8 . . . 1 ) d H a a /2 1 .5 d 2 d d -b o lt d ia m et er 0 .0 8 d d 2 d 2 d 3 K 1 an d K 2 B B L e 0 .0 6 d (0 .8 . . . 1 ) d (0 .8 . . . 1 ) d P u rs u an t to d (0 .3 5 . . . 0 ,4 ) a fo r o n est ep re d u ce r 0 .5 a fo r m u lt ist ep re d u ce r K 2 \u00fe d 1 \u00fe 5 m m ( 8 ",
            "48)], P is the power transmitted (in kW), u is the gear ratio, b is the facewidth (in mm), z1,2 is the number of teeth of the plastic gear of interest, v is the reference circle peripheral speed (in m/s), k is the exponent (k\u00bc 0.75 for PA; k\u00bc 0.4 for POM) and A is the cooling surface of a housing (in m2). Tables 4.11 and 4.12 contain calculation factors for Equations (4.46) and (4.47). The intermittence factor is equal to: f \u00bc 0:052 I0:64 \u00f04:48\u00de where I is the intermittence (relative operating time in 24 h). Symbol Unit Description Principal symbols and abbreviations a mm Centre distance A mm Dimension of housing (Figure 4.4) b mm Gear width; length of tooth C mm Dimension of housing (Figure 4.4) D mm Dimension of housing (Figure 4.4) d mm Diameter (without subscript, reference diameter) E Pa Gear elasticity modulus e mm Depth of dive in lubricant; dimension of housing (Figure 4.4) F N Force; load G Pa Shear modulus of lubricant H mm Dimension of housing (Figure 4.4) h mm Depth of teeth i Transmission ratio k W m2 K Overall heat transfer coefficient m mm Module n min 1 Rotational speed O mm Perimeter of inside surface of gear drive housing near the wheel gear P W Power PA \u2014 Material designation for polyamide POM \u2014 Material designation for polyoxymethylene p Pa Pressure Q \u2014 Gear accuracy grade _Q W Heat flow n min 1 Rotational speed r mm Radius s mm Housing wall thickness T Nm Torque t s Time of a single passing of oil through spray lubrication system u \u2014 Gear ratio m/s Tangential speed of contact point v m/s Peripheral speed x \u2014 Profile shift coefficient z \u2014 Number of teeth a W m2 K Convective heat transfer coefficient 8 Arbitrary circle (or point) pressure angle (without subscript, reference circle) b 8 Helix angle (without subscript, at reference cylinder) d mm Dimension of housing (Figure 4.4) e \u2014 Contact ratio h \u2014 Power efficiency Pa s Dynamic viscosity w Angle which the involved oil has to pass until mesh q K Temperature l W m K Coefficient of thermal conductivity m \u2014 Coefficient of friction n \u2014 Poisson number m2/s Kinematic viscosity r mm Radius of curvature v s 1 Angular speed c \u2014 Ratio of dimensions (b/mn) \u2014 Relative clearance in journal bearing 1 Pinion min Minimum value 2 Wheel gear n Normal plane I, II gear drive steps oil Oil a Tooth tip; addendum; axial direction P Permissible value b Base circle R Roughness C Pitch point r Radial c Cooler red Reduced e Outside t Transverse plane f Dedendum y Arbitrary point of tooth profile i Inside w Pitch circle L Losses Water lim Value of reference strength y Arbitrary point M Material e Contact ratio m Mean value S Sum max Maximum value AK m2 Surface of housing from which the heat is transferred vo m/s Peripheral speed of dived gear Ac m2 Surface of cooler tubes vto m/s Peripheral speed of tested gear bs \u2014 Width of thrust bearing segment vSY m/s Sum of local peripheral speeds bo mm Test gear width vSC m/s Sum of peripheral speeds in pitch point C1, C2 \u2014 Bath lubrication constants v \u2014 Speed parameter CSp \u2014 Factor of oil spraying vs m/s Speed of sprayed oil Cth \u2014 Heat correction factor wBt N/m Specific tooth load cp J/ (kg K) Specific worm capacity of oil at constant pressure Xab \u2014 Angle factor da2 mm Tip circle diameter of wheel gear YC \u2014 Depth of quenching factor eo mm Test gear depth of dive Yd \u2014 Design factor Ffr N Peripheral force of friction ae W m2 K Convective heat transfer coefficient from outer surface of housing wall to air fF \u2014 Correction factor for KV ai W m2 K Convective heat transfer coefficient from oil to inside surface of housing fK \u2014 Cooling factor an 8 Pressure angle of basic rack tooth profile fV \u2014 Viscosity factor ap m2/N Viscosity\u2013pressure coefficient Hh mm Height of housing a K 1 Viscosity\u2013temperature coefficient Hv \u2014 Factor of mesh power losses aw 8 Working pressure angle hC mm Oil film thickness in pitch point Dq K Oil temperature difference of input and output the cooler h \u2014 Relative tooth depth in pitch point d1, d2, d3 mm Dimensions of housing (Figure 4.4) lh mm Hydraulic length of gear drive housing e1 \u2014 Partial contact ratio of pinion K1, K2 mm Dimensions of housing (Figure 4.4) e2 \u2014 Partial contact ratio of wheel kc W m2 K Overall heat transfer coefficient from oil to cooling water ea \u2014 Transverse plane contact ratio Lth \u2014 Heat factor of lubricant eb \u2014 Overlap factor lB mm Journal bearing length eg \u2014 Sum contact ratio lh mm Hydraulic length of housing en \u2014 Contact ratio in normal plane ls mm Length of thrust bearing segment h0 Pa s Corrected dynamic viscosity mn mm Normal module hB \u2014 Efficiency of bearings PB W Power losses in bearings hoil Pa s Dynamic viscosity of oil at operating temperature PS W Power losses in seals hp Pa s Effective viscosity PL W Total power losses hZ \u2014 Mesh efficiency PZ W Mesh power losses hZ0 \u2014 Mesh efficiency at idle motion PZP W Mesh power losses under load qair C Environment temperature PZ0 W Power losses in idle motion qe C Temperature of outer side of housing pm Pa Mean pressure in bearing qeo K Difference qe \u2013 qair Ra mm Arithmetic mean roughness qoil C Operating temperature of oil _Q W Heat flow qoil;o K Difference of oil and environment temperature _Qc W Amount of heat flow to be carried away by cooling the oil in a cooler qw K Mean temperature of cooling water (Continued) _Qlim W Largest possible quantity of heat which can be transferred in environment qw1, qw2 C Input and output water temperature in cooler Qe l/min Quantity of sprayed lubricant ma \u2014 Axial sliding bearing coefficient of friction QeO l/min Quantity of sprayed lubricant of tested reducer mc \u2014 Coefficient of friction in pitch point So \u2014 Sommerfeld number mmz \u2014 Mean coefficient of friction in mesh Tfr Nm Moment of friction my Local coefficient of friction TH Nm Hydraulic moment of power losses noil m2/s Kinematic viscosity of oil Voil m3 Volume of oil in oil tank rCn mm Equivalent radius of meshed profiles _VP m3/s Oil pump flow roil kg/m3 Lubricant density vm m/s Peripheral speed at mean diameter of thrust bearing cb \u2014 Ratio b/d"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002122_978-3-319-00858-5_5-Figure5.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002122_978-3-319-00858-5_5-Figure5.12-1.png",
        "caption": "Fig. 5.12 Equilibrium of the nematic director in hybrid alignment conditions. a The bulk disclination line experiences an attractive force due to the opposite topological charges at the corners. b The defect settles at the corner in equilibrium state. c\u2013e Polarization micrographs showing cross-over of the disclination line from one side of the channel to the other. Multiple cross-overs were also observed (e). Scale bar: 50 \u00b5m",
        "texts": [
            " At corners formed by two PDMS walls, the alignment encounters a situation of nonconformity, which is resolved through two possible elastic deformations of the director field. In topological terms, the deformations correspond to regions of topological rank 1/4 at each corner. The net topological charge is nevertheless conserved through the development of a singular line defect, of opposite strength and rank 1/2, within the nematic bulk. However, these configurations do not correspond to the state of minimum free energy. As a consequence of a (logarithmic) attractive potential [24], defects of opposite topological charges approach each other (Fig. 5.12a). Hence, the semi-integer defect line settles at one of the corners (Fig. 5.12b). Using the relations of Dafermos [28] and Ericksen [29] for isotropic elasticity (K = 5.5 \u00d7 10\u221212 pN), the attractive force between the bulk disclination and each of the oppositely charged topological entity at a corner was estimated. The total energy Eelastic 12 and force of interaction Felastic 12 between two disclinations is given by: Eelastic 12 = 2\u03c0 K s1s2ln ( r12 rc ) (5.1) Felastic 12 = 2\u03c0 K ( s1s2 r12 ) (5.2) Here, s1 and s2 are the topological ranks of the interacting defects, and r12 is the separation between them",
            " In the present calculations, the interaction between the defects at the corners was neglected due to the large separation between them. The bulk disclination was thus found to be attracted with a force of Felastic \u2248 4 \u00b5N/m at a separation of 1 \u00b5m away from the corner. At the isotropic-to-nematic transition, the defect equilibrates with similar probability at either corner, resulting in a cross-over from one to the other side of the longitudinal confinement (l w), as observed clearly in polarization micrographs (Fig. 5.12c\u2013e). Figure 5.12b shows the resulting director field in the vicinity of the cross-over. The tendency of the defects to dwell near the walls can be overcome, however, by using electric fields [11]. In Chap. 6, we shall look at an alternative method to shift the director equilibrium\u2014by utilizing the viscous drag forces\u2014in presence of a flow field. The evolution of the static director configurations\u2014in absence of any external perturbation\u2014is of prime significance, as they register the initial conditions for the subsequent flow phenomena"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.38-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.38-1.png",
        "caption": "Figure 10.38 Flux concentration configurations: (a) configuration with the assistant magnets in series; (b) equivalent demagnetizing curve when the assistant magnets have the same mmf as that of the dominant magnet; (c) when the mmf of assistant magnets is more than that of the dominant magnet; and (d) when the mmf of assistant magnets is less than that of the dominant magnet",
        "texts": [
            "5, bt1 and ht1 are the width and height of the stator slots; bt2 and ht2 are the width and height of the rotor slots; ly1 and hy1 are the length and height of the stator yoke; ly2 and hy2 are the length and height of the rotor yoke; Af 1 and Af 1 are the cross-sectional areas of each magnetic bridge; and lf 1 and lf 2 are the lengths of each magnetic bridge, respectively. The operating point of magnets at rated load can also be solved by shifting the air-gap curve \u03b4 \u223c Fm by Fad to the left on the graph, where Fad is the armature mmf. Flux concentration configurations are often used in IPM machines to increase air-gap flux density as shown in Figure 10.38a. In flux concentration configurations, the assistant magnets are usually designed to have the same mmf as the dominant magnet. However, due to dimensional and other constraints, the assistant magnets may have a different mmf. There are three possibilities as shown in Figure 10.38b\u2013d. The Norton equivalent of all magnets can still be expressed by (10.76), with r and Fc the equivalent residual flux and equivalent mmf, and RM the equivalent reluctance. The equivalent r and Fc can be expressed as r = r1 + r2 (10.93) Fc = r r1Fc2 + r2Fc1 Fc1Fc2 (10.94) where r1 and Fc1 are the residual flux and mmf of the dominant magnet, and r2 and Fc2 are the residual flux and mmf of the assistant magnet. Sizing of magnets is one of the critical tasks of PM machine design. This section discusses the analytical methods to calculate the volume and size of magnets for PM motors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.5-1.png",
        "caption": "Figure 4.5 Three-dimensional design of cast housing of one-step reducer with cylindrical gears",
        "texts": [
            "6 A p p ro x im at e d im en si o n s o f ca st h o u si n g p u rs u an t to F ig u re 4 .4 a d d 1 C D h h 1 d 1 S ta n d ar d ce n tr e d is ta n ce 0 .0 3 a \u00fe5 m m (0 .8 . . . 1 ) d H a a /2 1 .5 d 2 d d -b o lt d ia m et er 0 .0 8 d d 2 d 2 d 3 K 1 an d K 2 B B L e 0 .0 6 d (0 .8 . . . 1 ) d (0 .8 . . . 1 ) d P u rs u an t to d (0 .3 5 . . . 0 ,4 ) a fo r o n est ep re d u ce r 0 .5 a fo r m u lt ist ep re d u ce r K 2 \u00fe d 1 \u00fe 5 m m ( 8 . . . 1 0 ) d 2 The three-dimensional design of a casted housing of a one-step reducer with cylindrical gears is presented in Figure 4.5. Welded Housings Awelded layout is applied in a single or low-series production. In principle, the thickness of the walls can be less than those of cast iron housings and covers. It usually takes 30% less than the thickness of cast housings and covers. Housing and housing cover are welded from elements prepared from a sheet. The layout of a welded housing of a coaxial two-stage reducer with welded cylindrical gears is presented in Figure 4.6. During operation, oil in the housing is heated and evaporates"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure17-1.png",
        "caption": "Fig. 17 Archimedean conic rack mechanism",
        "texts": [
            " 16) and therefore the operation should be split such that all the tips on one tooth flank are done and then all the tips on the opposite tooth flank are chamfered, which would improve cycle times. Collision Turn Table End Mill Tool Spindle Work Piece Fig. 14 \u201cAC\u201d type CnC machine limits for Toe/Heel chamfering\u2014small pitch cone angle bevel gear Tool Spindle Work Piece End Mill Turn Table Fig. 15 \u201cAC\u201d type CnC machine limits for Toe/Heel chamfering\u2014large pitch cone angle bevel gear The same comments apply when the pitch cone angle of a straight bevel gear is large, as is shown in Fig. 17. Tip chamfering can also be performed using the EM tool\u2019s tip which, in most cases, is a better approach since the turn table tilt is less and there are no risks of tool interference. However, the tool diameter becomes a limiting factor to avoid damaging the opposite tooth flank. When chamfering the Tip edges, the EM toolmust be tangent to the axial direction of the tip and angled to provide the desired chamfer (Fig. 18). Again, five unit vectors (left, Fig. 18) are required at any point along the tooth Tip edge: N : the local normal vector, 164 C"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003059_evs.2013.6915035-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003059_evs.2013.6915035-Figure1-1.png",
        "caption": "Figure 1: Detail from the stator, rotor and winding of our outer rotor and fractional slot per pole type motor",
        "texts": [
            " Keywords: efficiency, field-weakening, HEV (hybrid electric vehicle) There are some sophistically elaborated control methods in the literature to achieve the needed points of operation increase in speed and power regarding the available torque [1],[3],[6],[7]. Vector control method provides adequate possibilities for realise the tasks mentioned in the PM synchronous motor. Nowadays to apply an electrical driven for vehicle demands a PMSM having extended speed range possibility with high efficiency and a good torque-ampere ratio. With Infolytica software we had developed a 110kW PMSM for a city bus of 12 tons driven by battery, see Fig. 1. The nominal voltage, current, torque, speed are 650V, 240A, 1100 Nm and 1000 rpm respectively. The maximum torque is 2500 Nm under 900 rpm and the maximum speed is 2500 rpm with good features, it can be seen in Fig. 2-3-4-5. The outer type rotor has surface-mounted NeFeB magnets. For achieve a lower cogging torque we chose a fractional number of slots per pole. We have some experiences in this area throughout building some PMSM by lower power. EVS27 International Battery, Hybrid and Fuel Cell Electric Vehicle Symposium 2 We were studying to achieve the possible most appropriate rate of pole and slot numbers, magnet shape and thickness, its remanence, the airgap length, stack length, winding type, and the shape and measure to all detail of the slot"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.12-1.png",
        "caption": "Fig. 3.12",
        "texts": [
            " 3.12a,b). We assume that the wheel rolls (no slipping at point A). As the wheel moves, its center C travels a distance x along a straight horizontal line, point A moves to the location A\u2032, and the wheel undergoes a rotation of angle \u03d5. The arc length r\u03d5 and the distance x have to coincide: x = r\u03d5. Differentiating and using x\u0307 = vC and \u03d5\u0307 = \u03c9 yields vC = r \u03c9 . (3.10) The point of contact A with the ground momentarily has zero velocity (no slipping!): it is the instantaneous center of rotation \u03a0 (Fig. 3.12b). According to (3.9) and (3.10), the velocity of an arbitrary point P of the wheel is given by vP = rP \u03c9 = vC rP r . The velocity vector is perpendicular to the straight line \u03a0P . The maximum velocity is found at point B. With rB = 2 r we get vB = 2 vC . Let us now reconsider the motion of the rod in Fig. 3.7a. The given velocity vA is horizontal, the unknown velocity vB is vertical (see Fig. 3.12c). The rod momentarily rotates about the instantaneous center of rotation \u03a0 which is given by the point of intersection of the straight lines which are perpendicular to the velocities. The angular velocity \u03c9 of the rod can immediately be obtained from (3.9): vA = rA \u03c9 = l\u03c9 cos\u03d5 \u2192 \u03c9 = vA l cos\u03d5 . This leads to the velocity of point B: vB = rB \u03c9 = l\u03c9 sin \u03d5 = vA tan \u03d5 . As mentioned before, the instantaneous center of rotation \u03a0 is not a fixed point. Its location depends on the location of the rod. The locus of points which represent the instantaneous center of rotation in the space-fixed plane is called the centrode. In the present example it is a quarter-circle with radius l (Fig. 3.12c). E3.3 Example 3.3 The pulley system shown in Fig. 3.13a consists of two pulleys 1 and 2 and a disk 3 . The pulleys rotate with angular velocities \u03c91 and \u03c92, respectively. Determine the angular velocity of the disk and the velocity of its center C. Assume that the disk does not slip on the cable. Solution The velocities of the points A\u2032 and B\u2032 of the pulleys (Fig. 3.13b) are given by vA\u2032 = r1 \u03c91, vB\u2032 = r2 \u03c92 . 3.1 Kinematics 145 vA =vA\u2032 =r1\u03c91 C r1 r2 vB =vB\u2032 =r2\u03c92 A\u2032 B\u2032 vA\u2032 C vC \u03a0 rA BA \u03c91 \u03c92 r3 vB\u2032 1 3 2 a b \u03b1B \u03b1A Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003525_20120905-3-hr-2030.00102-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003525_20120905-3-hr-2030.00102-Figure1-1.png",
        "caption": "Figure 1: EMPS prototype and flexible coupling",
        "texts": [
            " Section 3 presents the usual identification method called IDIM-LS while section 4 presents the DIDIM method for a rigid robot. Section 5 presents the DIDIM method for a flexible joint manipulator. Section 6 is devoted to the experimental identification of one prismatic flexible joint manipulator. Then DIDIM method is 978-3-902823-11-3/12/$20.00 \u00a9 2012 IFAC 19 10.3182/20120905-3-HR-2030.00102 compared with IDIM-LS method and classical Output Error (OE) method. 2. MODELING OF A FLEXIBLE JOINT ROBOT 2.1 Experimental setup The EMPS is a high-precision linear Electro-Mechanical Positioning System (see Fig. 1). It is a standard configuration of a drive system for prismatic joint of robots or machine tools. Its main components are: - A Maxon DC motor equipped with an incremental encoder to make a PD motor position control. - A Star high-precision low-friction ball screw drive positioning unit. An incremental encoder measures the angular position of the screw. - A carriage which moves a payload in translation. These components are presented in Fig. 2. All variables and parameters are given in SI units on the load side"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001504_1-4020-3169-6_48-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001504_1-4020-3169-6_48-Figure2-1.png",
        "caption": "Fig. 2. 3-D FEM electromagnetic model, stator (on the left) and rotor (on the right)",
        "texts": [
            " By applying the magnetic forces to the structural model we can calculate the magnitude and frequencies of the resulting vibrations. Finally, the results of the structural analysis, represented by the velocities on the exterior surface of the electric motor, are used as an input for the BEM acoustic model. Magnetic Forces For the magnetic-force calculation in the investigated DC electric motor, two 3-D FEM models were built; the only difference between the models was in the rotor skewing angle. Fig. 2 shows the FEM model with the magnetic forces acting on the rotors and on the stator. To estimate the current loads in the commutation zones the assumption of a linear commutation was applied. By analyzing the magnetic forces acting on each magnet and on the rotor we found their amplitude and phase during the rotation for different motor currents of 0, 8, 16, 32 and 64A. Next, the variation in the resulting magnetic forces was decomposed by the discrete Fourier transform. Here, only the first five harmonics of the magnetic forces were calculated. As the rotor has twenty slots these harmonics are the 20th, 40th, 60th, 80th and 100th. Fig. 3 and Fig. 4 show the character of the most dominant component, i.e., the radial component, of the resulting magnetic force FR, see Fig. 2, acting on the magnet for both FEM models, with the skewed and non-skewed rotor. [N ] [N ] The calculation of the magnetic forces and the electromagnetic FEM model were verified by measuring the cogging torque on the motor with the non-skewed rotor. The reason for this is that the cogging torque of the motor with the non-skewed rotor is more distinctive and therefore more practical to measure. Fig. 5 shows the comparison between the numerically calculated and the measured data of the cogging torque during the rotation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003580_s1068798x13060099-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003580_s1068798x13060099-Figure1-1.png",
        "caption": "Fig. 1. Kinematic diagram of a doubly coupled mecha tronic system: X, Y, coordinate axes; the other notation is explained in the text.",
        "texts": [
            " As applications become more com plex, interest rises in multiply coupled mechatronic systems. In the present work, we consider a dynamic model of a doubly coupled mechatronic system with elastic ity and free play in the mechanical transmission, with a view to developing a tool for investigating the pro cesses in multiply coupled systems. For example, self oscillation due to the linearity in executive drives has not been sufficiently investigated for systems with mutual influence of the coordinates. In Fig. 1, we show the kinematic configuration of a doubly coupled mechatronic system, consisting of two rods. The system is characterized by two generalized coordinates \u03d5 and \u03c8 in kinematic pairs with collinear axes. We describe the system dynamics by means of sec ond order Lagrangian equations [1]. The kinetic energy is where J is the moment of inertia of the first body; \u03b8 is the moment of inertia of the second body; r is the length of the first rod; m is the mass of the second rod; s is the distance from the center of gravity of the sec ond rod to its point of suspension; and are gen eralized angular velocities; \u03b2 = \u03c8 \u2013 \u03d5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003656_s0967091213080081-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003656_s0967091213080081-Figure2-1.png",
        "caption": "Fig. 2. Rolling of strip inclined toward the upper working roller (a) and wear of the rough surface of the upper (1) and lower (2) working rollers (b). Initial roughness Ra = 6\u20136.3 \u03bcm. Cell 1 of the 1680 continuous cold rolling mill at Zaporozhstal\u2019 metal lurgical works.",
        "texts": [
            " In practice, however, the asym metry may be associated with a new factor. Thus, in a continuous cold rolling mill, the wear at the upper and lower working rollers is often different. That is due to inclination of the strip toward one of the rollers [5, 9, 10]. STEEL IN TRANSLATION Vol. 43 No. 8 2013 COMPLEX ROLLING OF THIN STRIP 517 Inclination of the strip toward one of the rollers changes the conditions of lubricant capture. Experi ments [11] indicate that the thickness of the lubricant layer is greater at the side of the roller to which the strip is inclined (Fig. 2a). As follows from Fig. 2a, the input angle of the lubricant to the deformation zone is less at the upper roller. Consequently, the lubricant layer is thicker at the upper surface of the strip than at the lower roller, while the frictional coefficient is less. Nevertheless, the reduction \u0394hU at the upper roller is less [5, 9]. With decrease in the reduction, the forces will act differently at the upper and lower rollers. The equilibrium of the forces in the vertical plane takes the form (Fig. 2a) PU + Plu.U = PL + Plu.L. (7) where PU and PL are the rolling forces in the geometric deformation zone at the upper and lower rollers, respectively; Plu.U and Plu.L are the forces exerted on the oil by the upper and lower rollers, respectively. Since the length of the lubricant wedge at the upper roller is greater, Plu.U > Plu.L. Then it follows from Eq. (7) that PU < PL. That is possible if the reduction \u0394hU and contact length lU at the upper roller are less than the values \u0394hL and lL at the lower roller",
            " In those conditions, the lower surface of the strip is in contact with the lower roller until it leaves the deformation zone. On account of the strip creep in the deformation zone, its speed before enter ing the deformation zone is significantly less than the azimuthal velocity of the rollers vro/vdr \u2248 1 + Scr. (8) On account of the considerable roller slip at the surface of the strip, there is constant additional wear of the microprojections. Therefore, the surface of the lower roller will always be less rough (Fig. 2b). In cells 2\u20134, the strip deviates toward the upper roller, and hence the wear at the upper working rollers is greater. That impairs the overall performance of the rollers in the finishing cell (cell 4), since intense wear of the upper working rollers necessitates replacement of the whole set. To maintain the roughness of the upper working roller in cell 4, it is expedient to ensure special regular surface relief of the upper working roller in cell 3, rather than using a ground roller [5]. The greater roughness of the initial strip ahead of the finishing cell ensures that more lubricant is supplied to the deformation zone, and thereby the wear is reduced"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001531_s00707-013-0995-y-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001531_s00707-013-0995-y-Figure6-1.png",
        "caption": "Fig. 6 General subsystem: top view on the motor axis. rpi \u2208 IR3 (as well as system i) viewed as a projection onto the (x\u2013y)-plane",
        "texts": [
            " . \u00b7 \u00b7 \u00b7 ... . . . ... FT N \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u239b \u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239d M1y\u0308R1 + G1y\u0307R1 \u2212 Q1 ... Mi y\u0308Ri + Gi y\u0307Ri \u2212 Qi ... ... MN y\u0308RN + GN y\u0307RN \u2212 QN \u239e \u239f\u239f\u239f\u239f\u239f\u239f\u239f\u239f\u23a0 = 0. (44) Because D does not undergo an integration by parts, the (dynamic) boundary conditions do not arise but are implicitly fulfilled. Of course, the shape functions \u03a6(x) in Eq. (41) have to fulfill geometrical demands (and to form a complete set of functions). Considering Eq. (44) also holds for more realistic subsystems as for instance depicted in Fig. 6 [7]. Here, the reference frame R moves with vR \u2208 IR3 and \u03c9R \u2208 IR3. However, because the motor angular velocity M is always uniaxial, it will be suitable to split the angular velocity into \u03c9R = \u03c9F + Mr e3 (index F : \u201cF\u00fchrung\u201d (guidance), index r : relative). The gear output angle \u03b3M (\u2192 iG \u03b3\u0307M = Mr , iG : gear ratio) and the arm angle \u03b3A are, consequently, also uniaxial. However, the vector that describes the position of a beam element rT = [x, v(x, t), w(x, t)] as well as the corresponding angles \u03d5T = [\u03d1(x, t), \u2212w\u2032(x, t), v\u2032(x, t)] is three-dimensional",
            " For a rigid body, the linear momentum is p = m vc and the angular momentum L = J \u03c9c, respectively, (index c: mass center). If calculated in a reference frame representation, not only \u03c9c is non-integrable (as in any representation), but also vc. However, the most practicable mechanical modeling refers to an arbitrary reference frame, where the equations of motion are most easily obtained by the Projection Equation N\u2211 i=1 [( \u2202vc \u2202 s\u0307 )T ( \u2202\u03c9c \u2202 s\u0307 )T ] i ( (p\u0307 + \u03c9\u0303R p \u2212 fe) (L\u0307 + \u03c9\u0303R L \u2212 Me) ) i = 0, (45) (fe, Me: impressed forces and torques, e.g., gravitation and gear spring, see Fig. 6). The minimal velocities s\u0307 = H(q)q\u0307 are adequately chosen. The Jacobians (\u2202vc/\u2202 s\u0307) and (\u2202\u03c9c/\u2202 s\u0307) are nothing but the coefficient matrices of vc and \u03c9c w.r.t. s\u0307; they are obtained as a by-product when calculating vc and \u03c9c (there is no additional calculation needed!). Equation (45) overcomes all the difficulties that possibly may arise with nonholonomic variables [6]. Moreover, non-holonomic constraints, if occur, may \u00e0 priori be considered (which is not allowed for the analytical procedures [15])",
            " Elimination means to write down the momentum equations and then to eliminate line by line the (generalized) constraint forces (somehow comparable to D\u2019Alembert\u2019s principle [9]). Projection, on the other hand, means to project the momentum equations into the unconstrained (\u201cminimal\u201d) space (as an application of Lagrange\u2019s principle (virtual work) [10]). Its mathematical evaluation then leads to the Gauss-form (44) where simultaneously subsystems enter the calculation. The simplest subsystem is the single (here: elastic) body itself but in general, it characterizes an assembly group (like in Fig. 6 for instance). This foregoing yields clear-cut advantages not only concerning modeling aspects but also to the clear structured representation of the basic dynamics. (Its \u201corder-n-evaluation\u201d needs three steps: kinematics from bottom to top, kinetics from top to bottom, resolution of minimal accelerations from bottom to top; for details see, e.g., [7,14]). Up to here, the consideration is based on the use of a Ritz ansatz that leads to ordinary differential equations. But during all the time, the question frequently came up how the corresponding PDEs look like"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.48-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.48-1.png",
        "caption": "Figure 12.48 A 2R planar manipulator carrying a load at the end point.",
        "texts": [
            " The forward Newton\u2013Euler equations of motion allow us to start from a known action force system (1F0, 1M0), where the base link applies to the link (1), and calculate the action force of the next link. Therefore, analyzing the links of a multibody one by one, we end up with the force system where the end effector applies to the environment. Using the forward- or backward-recursive Newton\u2013Euler equations of motion depends on the measurement and sensory system of the multibody. Example 763 Recursive Dynamics of a 2R Planar Manipulator Consider the 2R planar manipulator shown in Figure 12.48. The manipulator is carrying a force system at the end point. We use this manipulator to show how to develop the dynamic equations for a serial multibody. The backward-recursive Newton\u2013Euler equations of motion for the first link are 1F0 = 1F1 \u2212 \u2211 1Fe1 + m1 1 0a1 = 1F1 \u2212 m1 1g + m1 1 0a1 (12.754) 1M0 = 1M1 \u2212 \u2211 1Me1 \u2212 (1d0 \u2212 1r1 )\u00d7 1F0 + (1d1 \u2212 1r1 )\u00d7 1F1 + 1I1 1 0\u03b11 + 1 0\u03c91 \u00d7 1I1 1 0\u03c91 = 1M1 \u2212 1n1 \u00d7 1F0 + 1m1 \u00d7 1F1 + 1I1 1 0\u03b11 + 1 0\u03c91 \u00d7 1I1 1 0\u03c91 (12.755) and the backward-recursive equations of motion for the second link are 2F1 = 2F2 \u2212 \u2211 2Fe2 + m2 2 0a2 = \u2212m2 2g \u2212 2Fe + m2 2 0a2 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003054_ijtc2012-61041-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003054_ijtc2012-61041-Figure2-1.png",
        "caption": "Fig 2 Four tilting pad cage, current design.",
        "texts": [
            " The analysis to determine the range of design preload as a function of oil and housing temperature has been automated in an Excel file. This allows a quick determination of the min and max expected clearance and preload for the cold build and also both no-load and loaded operation. The original design had a nominal 3 mil on diameter clearance with a preload range of 0.35 to 0.57. An offset of 60% was selected to increase the load capacity and reduce the pad operating film temperature. The new modified bearing cage and pads are shown in Fig 2. With a turbine imbalance of 0.72 g-mm (0.001 oz-inch), the predicted response at the compressor end bearing station 8, and turbine bearing station 9, are shown in Fig 3 for the 1 mil on diameter clearance condition. The first and second modes are predicted to be at 41 and 66 kRPM respectively, from this forced response analysis. The typical spectrum content on the stock bearings has been shown in previous papers to have the first mode in the 12- 20 kcpm frequency range, even at the engine idle speed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002978_detc2011-48760-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002978_detc2011-48760-Figure1-1.png",
        "caption": "Figure 1. Pitch cones of straight bevel gears.",
        "texts": [
            " In addition, the tooth contact pattern of the manufactured largesized straight bevel gears was compared with those of tooth contact analysis. As a result, there was good agreement. The basic design parameters of straight bevel gears are the normal module Mn; the number of teeth of pinion and gear zp and zg, respectively; the shaft angle  and the pressure angle . The gear ratio is given, in the same manner as that of other types of gears, as follows: p g z z i  (1) The pitch surfaces of bevel gears are cones as shown in Fig. 1. Therefore, the following equation yields: i d d g p 0 0  (2) where dp0 and dg0 are the pitch circle diameters (PCD) of the pinion and gear, respectively. The pitch cone angles of the gear and pinion are represented by 00 1 0 cos1 sin tan gp g i        (3) The outer cone distance is determined using Eq. (3) as follows: 0 0 sin2 g g e d R   (4) Therefore, the mean and inner cone distances are as follows, respectively: bRR b RR ei em   2 (5) where b is the face width. The normal module is defined as that in the center of the tooth surface in this paper although it is usually chosen as that in the heel side"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.8-1.png",
        "caption": "Fig. 3.8 Quadrifilar pendulum. (a) Side view. (b) Top view. (c) Top view, plate rotated through angle",
        "texts": [
            ", the point mass, m, moves on a horizontal small circle (Fig. 2.12d). In this special case a rotating simple rod pendulum can be called a circular pendulum, or a conical pendulum. The time of swing for one complete revolution of the small circle is given by Eq. 2.37. From this equation (Lineham 1914), the vertical distance of the point mass, m, below the pivot, lV (Fig. 3.7) is lV D g !2 (3.9) where g is the acceleration due to gravity, and ! is the angular velocity. A quadrifilar pendulum is shown in Fig. 3.8. It consists of a uniform rigid square plate, diagonal, 2c, mass, M , suspended by four vertical inextensible massless strings, length, l . In effect, it is four coupled simple string pendulums (Sect. 2.4). Provided that the strings remain taut a quadrifilar pendulum has three degrees of freedom and, because the strings are of equal length, the plate remains horizontal. If the rigid plate does not rotate then there are two degrees of freedom, and the motion of each point mass, m, is in phase and identical and also, in general, identical to the motion of the point mass, m, of a simple string pendulum (Sect. 2.4). Exceptions are due to inference between components of the quadrifilar pendulum. For small amplitudes the time of swing, T , is given by Eq. 2.13 where g is the acceleration due to gravity. The third degree of freedom means that a torsional mode of oscillation is possible. In this torsional mode of oscillation the uniform rigid square plate rotates about a vertical axis through its centre. If the plate is turned through a small angle, , about its centre (Fig. 3.8c) then the lower end of each string moves through an arc of length approximately c , where 2c is a diagonal of the square and is in radians. The tensile force in each string is approximately 1=4Mg, where M is the mass of the plate, and g is the acceleration due to gravity. The horizontal component of the tensile force is approximately 1=4Mgc /l , and the total restoring couple is Mgc2 =l . The corresponding differential equation is (Lamb 1923) M 2 d2 dt2 D Mg c2 l (3.10) where is the radius of gyration of the square, and t is time"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001268_3-540-34319-9_8-Figure8.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001268_3-540-34319-9_8-Figure8.4-1.png",
        "caption": "Figure 8.4: Mecanum principle, vector decomposition",
        "texts": [
            " For the construction of a robot with four Mecanum wheels, two left-handed wheels (rollers at +45\u00b0 to the wheel axis) and two right-handed wheels (rollers at \u201345\u00b0 to the wheel axis) are required (see Figure 8.3). Omni-Directional Drive Although the rollers are freely rotating, this does not mean the robot is spinning its wheels and not moving. This would only be the case if the rollers were placed parallel to the wheel axis. However, our Mecanum wheels have the rollers placed at an angle (45\u00b0 in Figure 8.1). Looking at an individual wheel (Figure 8.4, view from the bottom through a \u201cglass floor\u201d), the force generated by the wheel rotation acts on the ground through the one roller that has ground contact. At this roller, the force can be split in a vector parallel to the roller axis and a vector perpendicular to the roller axis. The force perpendicular to the roller axis will result in a small roller rotation, while the force parallel to the roller axis will exert a force on the wheel and thereby on the vehicle. Since Mecanum wheels do not appear individually, but e.g. in a four wheel assembly, the resulting wheel forces at 45\u00b0 from each wheel have to be combined to determine the overall vehicle motion. If the two wheels shown in Figure 8.4 are the robot\u2019s front wheels and both are rotated forward, then each of the two resulting 45\u00b0 force vectors can be split into a forward and a sideways force. The two forward forces add up, while the two sideways forces (one to the left and one to the right) cancel each other out. 8.2 Omni-Directional Drive Figure 8.5, left, shows the situation for the full robot with four independently driven Mecanum wheels. In the same situation as before, i.e. all four wheels being driven forward, we now have four vectors pointing forward that are added up and four vectors pointing sideways, two to the left and two to the right, that cancel each other out"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002491_imece2013-62177-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002491_imece2013-62177-Figure9-1.png",
        "caption": "Figure 9. a) Coefficients calculation along cycle C13, b) along cycle C14, c) along cycle C15",
        "texts": [
            "org/about-asme/terms-of-use           +\u2212\u2212 +\u2212+\u2212 +\u2212\u2212+\u2212 = == 100101000 010010101 001011011 C 151431121110987 15 14 13 36 C C C pairsgearcpairsturningt 4847644444 844444 76 (46) The independent set of equations for relative angular velocities, shown in (47)           =                     \u22c5           \u22c5\u2212\u22c5\u2212 \u22c5\u2212\u22c5+\u22c5\u2212 \u22c5\u2212\u22c5\u2212\u22c5+\u22c5\u2212 0 0 0 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 P10P1000 0P10P10P1 0P1P10P1P1 12 11 10 9 8 7 12,1510,15 11,149.147,14 11,1310,138,137,13 & & & & & & (47) The coefficients, from \u201cmoment\u201d equations with respect to gear pairs, shown in Fig. 9a, Fig. 9b, and Fig. 9c: 511,13510,1328,1327,13 d0.5P;d0.5P;d0.5P;d0.5P =\u2212=\u2212=\u2212= 411,1439,1437,14 d0.5P;d0.5P;d0.5P === 612,15 ' 410,15 d0.5P;d0.5P \u2212== The equations\u2019 coefficients as function of teeth ratios, (48) 6 ' 4 6 ' 4 15 4 3 4 3 14 5 2 5 2 N N d d ; N N d d ; N N d d ====== iii13           =                     \u22c5           \u2212 \u2212\u2212 \u2212\u2212 0 0 0 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 10i000 010i0i 0110ii 12 11 10 9 8 7 51 4141 1313 & & & & & & (48) The above system is solved for relative velocities as a function of input velocities, the solution is shown in (49) ( )           \u22c5\u22c5+\u22c5\u22c5+\u22c5+\u2212 \u22c5+\u22c5\u2212 \u22c5+\u22c5+\u22c5+\u2212 =           =           987 9714 98714 5/6 1/4 4/5 12 11 10 \u03b8ii\u03b8ii\u03b8iii \u03b8i\u03b8i \u03b8i\u03b8i\u03b8)i(i \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 &&& && &&& & & & & & & 15141513141315 14 141313 (49) The mechanism\u2019s absolute angular velocities 0 \u03c9 as a function of relative velocities are shown in (50)                     \u22c5                     \u22c5\u2212\u22c5\u2212\u22c5\u2212\u22c5\u2212 \u22c5\u2212\u22c5\u2212\u22c5\u2212 \u22c5\u2212\u22c5\u2212 \u22c5\u2212 \u22c5\u2212 \u22c5\u2212 \u2212=                     12 11 10 9 8 7 0 12 0 11 0 10 0 7 0 11 0 10 0 7 0 11 0 7 0 9 0 8 0 7 6 5 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 111001 011001 010001 000100 000010 000001 & & & & & & uuuu uuu uu u u u \u03c9 \u03c9 \u03c9 \u03c9 \u03c9 \u03c9 0 0 0 4 0 3 0 2 0 1 (50) With the unit vectors of turning pairs as shown in (51): ( ) ( ) ( )T TT 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure26-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure26-1.png",
        "caption": "Fig. 26 Toe chamfering with EM\u2014spiral bevel pinion",
        "texts": [
            " This approach is therefore not acceptable and the EM tool tip is to be used which, as shown in Fig. 24, causes no issue on either flank. Figure 25 shows the Tip edges of a small spiral bevel pinion chamfered using an EM. The chamfers are almost small enough to confuse with deburring. When chamfering tooth Toe and Heel edges, other considerations arise. For one, when chamfering the Toe of a pinion with a small pitch cone angle, again the turn table angle is likely to exceed the machine\u2019s limit and collide with the tool spindle (left, Fig. 26). Beyond this, even a small diameter EndMill is likely to interfere with the concave tooth flank when chamfering the bottom of the tooth (right, Fig. 26) and therefore either the Pivot Angle must be reduced, or else this solution becomes unacceptable. Tool spindle to turn table collision is not likely to occur at Heel (left, Fig. 27); and if the Pivot Angle is correctly chosen, tool interference with the tooth flank can be avoided (right, Fig. 27). The Ball Mill tool (BM), thanks to its spherical end, can be fitted in places where an End Mill tool would not do an acceptable job. Consider for example the spiral bevel pinion shown in Fig. 28, left"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003318_icist.2012.6221607-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003318_icist.2012.6221607-Figure5-1.png",
        "caption": "Fig. 5. The setup tools of experiment.",
        "texts": [
            " From the experiment simulation by adjusting the jetting target out from center show that group 2 category happen when the jetting target has moved in range 20 to 30 micrometer from center of connector to consecutive connector. The more off-center distance will create the bridging size in the group 3 and group 4 categories. The experiment results are shown in figure 4. B. Pin Layer Reversal and Reliability Test The experiment is setup to find the maximum temperature limit on TuMR reader head by using the hot air blow to TuMR reader head which mount on Head Gimbals Assembly and put the thermocouple on head to measure the temperature. The setup tools of the experiment are show in figure 5. This setup equipment is attached on the quasi tester machine which uses to determine the transfer curve of reader head the experiment. The flowchart of experiment is shown in figure 6. The HGA are load to fixture which is attached to quasi tester. The hot air tip blows from 20 degree angle to reader with 10 millimeters distance. The flow rate of hot air is adjusted to 1.5 feet per second through three millimeters diameter tip. The temperature will be measure and feed back by thermocouple type K which is installed above magnetic head about 500 micrometers"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.32-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.32-1.png",
        "caption": "Fig. 2.32 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PPaPR (a) and 4PPaRP (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology P\\Pa||P\\\\R (a) and P\\Pa\\kR\\kP (b)",
        "texts": [
            " 4PPRPa (Fig. 2.28b) P\\P||R\\Pa (Fig. 2.21f) The third joints of the four limbs have parallel axes 7. 4PRPPa (Fig. 2.29a) P\\R||P\\Pa (Fig. 2.21g) Idem No. 5 8. 4PRPaP (Fig. 2.29b) P\\R\\Pa \\kP (Fig. 2.21h) Idem No. 5 9. 4PPPaR (Fig. 2.30a) P\\P\\kPa\\kR (Fig. 2.21i) Idem No. 4 10. 4PPPaR (Fig. 2.30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003141_amm.86.908-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003141_amm.86.908-Figure1-1.png",
        "caption": "Fig. 1 Structure of the spline pair in wind turbine gearbox",
        "texts": [
            " The practical thermal inserting deformation on the spline was measured. And the results indicate that the measuring deformations are in good consistency with the deformation calculated by finite element analysis method. The spline deformation and tooth relieving calculation method can provide a reference for the tooth relieving of the spline on long structure sun gear shaft in 1.5MW wind turbine gearbox. In megawatt stage wind turbine gearbox, the long sun gear is a typical load sharing structure used in the planet stage. As shown in Fig.1, there is a spline pair between sun gear and the next parallel stage to transmit torque. Because of the big interference fit joint between the ring gear and axis with inner spline, the backlash and the gear profile will change during the thermal inserting process. And there will has torsion deformation caused by the high working torque acted on the spline. So, the backlash and the tooth profile of the spline need to be modified in order to ensure the load sharing behavior and transmission stationarity of the planetary gear stage, and to ensure the meshing strength of the spline"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003569_esda2012-82142-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003569_esda2012-82142-Figure1-1.png",
        "caption": "FIGURE 1: THE PHYSICAL MODEL OF THE GEAR PAIR",
        "texts": [
            " Moreover the technique is particularly suitable to detect discontinuities, such as jumps or impacts, and can help in the definition of indices and metrics suitable for rattle quantization. As regard to the theoretical model it is possible to obtain the relative gear motion described by a single degree-of-freedom lumped parameter model, considering the driven gear forced to vibrate by a motion imposed on the driving gear [3]. Motion is considered along the line of contacts. Model keeps in count the squeeze actions due to the oil interposed between the gears teeth. Figure 1 shows the physical model of the gear pair where the subscript 1 indicates the driving gear. The mathematical model is then: \u23aa \u23a9 \u23aa \u23a8 \u23a7 <<\u2212 \u2212\u2264 \u2265 = \u2212=++ bxhbxS hbxxxS bxxXK xxF XmfxxFxm r )( if )( if ,)( if ,)( ),( ,),( minmax min with & && &&&&& (1) where m is the mass of the driven gear, x = r2\u03b82 - r1\u03b81 is the relative displacement of the gears evaluated along the line of contacts, X = r1\u03b81 is the absolute motion of the driving gear, b is the backlash, and hmin represents the film thickness of the oil layer that remains adsorbed to the tooth surface during the contact phase. Moreover to keep in count the resistant actions arisen in the bearings supporting each unloaded gear idle on its shaft, a constant drag torque applied to the driven gear has been here considered, corresponding in eq. (1) to a constant force fr at the pitch point, calculated for stationary operating conditions in terms of loads, speed and oil parameters [3,4] During the contact phases, a non-linear elastic force acts on the driven gear (Fig. 1: case a, c). Such elastic force is given by the sum of the contributes due to the n teeth pairs that are in contact at the same time. Each teeth pair contributes with a stiffness term [16]: \u239f \u239f \u23a0 \u239e \u239c \u239c \u239d \u239b \u2212 = 3 z\u03b1 z api X\u03b51.125 2/X\u03b5X Cexpk(X)K (2) where \u03b5 and \u03b5\u03b1 are the total contact ratio and the transverse contact ratio respectively and Xz is the transverse base pitch. For a complete meshing cycle, X starts from 0 and ends at \u03b5 Xz. Finally, kp indicates the stiffness at the pitch point that depends on the tooth parameters along with the Ca coefficient. The elastic force is consequently the sum of n periodic functions shifted each other of a transverse base pitch Xz. The contacts can occur either on the driving or on the driven side of the tooth. During the approach phase when the teeth are separated through an oil layer (Fig. 1: case b), the oil squeeze effect gives a non-linear damping force ( )xxS & , with: ( ) ( ) ( ) ( )222/3 222 23 arctan 3 xRax xR axRaxRxRaa ZRxS + \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 \u239f \u23a0 \u239e \u239c \u239d \u239b++\u2212 \u2212= \u03bc (3) in which \u03bc is the absolute viscosity, Z indicates the axial width of the gear pair, R is the relative curvature radius of the teeth and a denotes the semi-length of the oil film along the tooth profile related to the lubrication conditions [3]. In the (1) Smax = S(hmin) represents the saturation value of the squeeze damping coefficient"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.17-1.png",
        "caption": "Figure 12.17 A suspended rigid body with a fixed point at geometric center.",
        "texts": [
            " In a spherical rigid body, every line that goes through the mass center is a principal axis. So, the dynamics of the body is independent of its attitude. Euler equations are dynamically coupled by the variable \u03c9i that appears in equations for Mj , j = i. They are also parametrically coupled by the difference of the principal mass moments. When the principal mass moments Ii, i = 1, 2, 3, are equal, the equations uncoupled as Equations (12.313)\u2013(12.315). The uncoupled equations are the simplest differential equations with solutions (12.316)\u2013(12.318). The suspended body in Figure 12.17 is a symmetric rigid body with a fixed point at its mass center if its width, height, and length are equal. 12.4.3 Axisymmetric Torque-Free Rigid Body A rigid body is called axisymmetric or axially symmetric if I1 = I2 = I3 (12.319) The torque-free Euler equations of an axisymmetric rigid body are I1\u03c9\u03071 \u2212 (I1 \u2212 I3) \u03c92\u03c93 = 0 (12.320) I1\u03c9\u03072 \u2212 (I3 \u2212 I1) \u03c93\u03c91 = 0 (12.321) I3\u03c9\u03073 = 0 (12.322) It provides a harmonic solution for angular velocity B G\u03c9B : \u03c91(t) = \u03c9\u03b8 cos t (12.323) \u03c92(t) = \u03c9\u03b8 sin t (12.324) \u03c93 = I1 I3 \u2212 I1 (12",
            " The solutions of cases 1 and 2 approach each other when D \u2192 I2, and we have k = 1 (12.458) \u03c91 = s1 \u221a 2K (I2 \u2212 I3) I1 (I1 \u2212 I3) 1 cosh \u03c4 (12.459) \u03c92 = s2 \u221a 2K I2 tanh \u03c4 (12.460) \u03c93 = s3 \u221a 2K (I1 \u2212 I2) I3 (I1 \u2212 I3) 1 cosh \u03c4 (12.461) If \u03c4 \u2192 \u221e, then \u03c92 \u2192 0, \u03c93 \u2192 0, and \u03c92 \u2192 s2 \u221a 2K/I2, which is a permanent rotation about the second principal axis. Torque-free rigid body equations are applied to celestial bodies which are almost free of external torques. They also apply on a body which is suspended frictionless such that its mass center is stationary. Figure 12.17 illustrates an example of such a suspended rigid body. Example 745 Orientation of B in G We are able to determine the time behavior of the components of B G\u03c9B . To determine the orientation of B in G for a given B G\u03c9B , we should select a set of generalized orientation coordinates and use the associated method of Example 718. Let us select the Euler angles as the generalized coordinates. Using the associated transformation matrix BRG of Equation (4.142), BRG = Rz,\u03c8Rx,\u03b8Rz,\u03d5 = \u23a1 \u23a2 \u23a3 c\u03d5c\u03c8 \u2212 c\u03b8s\u03d5s\u03c8 c\u03c8s\u03d5 + c\u03b8c\u03d5s\u03c8 s\u03b8s\u03c8 \u2212c\u03d5s\u03c8 \u2212 c\u03b8c\u03c8s\u03d5 \u2212s\u03d5s\u03c8 + c\u03b8c\u03d5c\u03c8 s\u03b8c\u03c8 s\u03b8s\u03d5 \u2212c\u03d5s\u03b8 c\u03b8 \u23a4 \u23a5 \u23a6 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003356_amm.86.871-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003356_amm.86.871-Figure1-1.png",
        "caption": "Fig. 1 Schematic of processing arc cylinder-gears Fig. 2 Tooth profile of linear edge of blade fillet",
        "texts": [
            " The paper uses generating method to process arc cylinder-gear, blade fillet and top tooth edge to create tooth root fillet of cylindrical gear and set up radius vector equation between working tooth surface and tooth root fillet of arc cylinder-gear, MATLAB programming to achieve accurate modeling of liner tooth surface and fillet of arc cylinder-gear, which can provide theoretic basis for analysis on tooth contact stress, tooth bending stress and kinetic. Use imaginary cutter-gear to conjugate and mesh with arc cylinder-gear and arc tooth surface gear simultaneously, to ensure the machined arc cylinder-gear can correctly mesh with arc tooth surface gear. The cutter tooth is equivalent to one tooth of imaginary cutter-gear, as shown in Fig. 1 \u2013 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: 152.14.136.77, NCSU North Carolina State University, Raleigh, USA-03/05/15,14:08:32) schematic of processing arc cylinder-gear. The movement of cutter consists of two parts, one is cutting movement, making it rotate on its own shaft at a speed of \u03c9f, which forms the main movement of cutting gear blank with cutter tooth, and has nothing to do with the imaginary cutter-gear and generation of gear blank"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002278_2013-01-1491-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002278_2013-01-1491-Figure10-1.png",
        "caption": "Figure 10. 6 speed manual transmission Romax Model",
        "texts": [
            "435mm, which is added into run out of the gear due to manufacturing error. Also increase in run out at the dog teeth decreases the shift quality such as, hard shifting and gear jump out. A 6 speed manual transmission of medium duty commercial vehicle has been studied by using Romax designer software and analyzes the effect of radial gap between the gear and the needle roller bearing on contact stress, transmission error and misalignment. A Romax model of 6 speed manual transmission has been shown in Figure 10. The Romax model of transmission has been run and analyze for the 2nd speed at defined load case. The transmission error, mesh stiffness and the contact pattern with the designed gear macro geometry has been analyzed. The micro geometry parameters like lead crowning and tip relief and, radial gap and run out (calculated) are not considered for this analysis. The load distribution during meshing of the gear tooth is not uniform and edge loading is observed as shown in Figure 11. Also, the transmission error or misalignment of the gear tooth i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002422_amm.271-272.1356-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002422_amm.271-272.1356-Figure2-1.png",
        "caption": "Fig. 2 The virtual prototype of positioning stage",
        "texts": [
            " In its modular named \u201cMechanism\u201d, kinematic and dynamic analysis and simulations can be carried out in order to look into the virtual movement of the mechanism, thus to get the kinematics parameters of the object throughout the process. The friction model of the mass-spring-dampness system has been built, from which the parameters causing stick-slip are acquired. So as to support the theoretical analysis, the virtual prototype of the positioning stage is established in the Pro/E three-dimensional design software, as is shown in fig 2. In the module named \u201cMechanism\u201d, objects on the guide, spring and damper between them are contained given separate properties. The driving object has initial velocity, static and kinetic friction also exists. Defining the velocity of the driven object as measuring object, data of the velocity will be obtained after simulation of the model in the module, meanwhile, reflecting the change of stick slip phenomenon. Influence of stiffness. As is shown in fig 3, the stick-slip phenomenon changing with different stiffness is presented"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure3-1.png",
        "caption": "Figure 3. Combination of Steel Press Center structure and Plastics.",
        "texts": [
            " A composite material suspension arm is developed with the same strength and stiffness compared with forged steel upper arm. And durability verification is proved the part and vehicle durability test based on Hyundai Motors specifications. The basic structure of composites suspension arm fixed to mold from steel press and plastic is injected into mold. However it is easily separated if the coefficient of friction, surface and material properties in the structure are different from other structure. Steel press center structure is a strong combination because burring fix vertical direction as shown in Figure 3. Steel press center structure is a strong combination because flange is fixed horizontal direction as shown in Figure 3. So the structure of steel and plastic were fixed in any direction. In general, ball joints that deliver the force from the tires are important parts with higher strength and durability. Also because ball stud often causes oscillation and rotation, ball joints bearings must be strong structure to withstand the steady force. Ball stud and steel pipes are fixed to the mold and plastic is injected into the mold in order to increase the strength and stiffness for ball joint part such as a Figure 4. For ball stud in order to smooth rotate and oscillate, plastic and bearings should be well fixed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003514_amm.101-102.224-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003514_amm.101-102.224-Figure1-1.png",
        "caption": "Fig. 1 Mechanism diagrams of rotational swash plate pulse CVT",
        "texts": [
            " Rotational swash plate pulse CVT describes in this article is a new design. It has the advantages which are not possessed by the existing pulse CVT. Its transmission component is composed of swash plate, guide rod and helical gear. It overcomes the shortcomings that inertia force or inertia moment are difficult to be balanced in the existing connecting rod pulse CVT or swing link pulse CVT. Its pulsation is low and the value of the pulsation does not vary with the change of the transmission ratio. Mechanism Diagram. As shown in Fig.1, \"1\" is axis of input and it is driving member. \"2\" is slider and it connects with axis of input in the form of sliding pair.\"3\" is a disk and its obliquity can be adjusted. Disk connects with axis of input at C point in the form of revolute joints (hinge). \"4\" is connecting rod and it connects with slider \"2\" in the form of hinge at the end point A. On the other end point B, it connects with swash plate \"3\"in the form of hinge. When the input shaft \"1\" rotates, the slider \"2\", swash plate \"3\" and the connecting rod \"4\" are driven to rotate at the same speed",
            " The so-called axis of helical gears meshing means: active helical gear do axial movement, driven helical gear do fixed axis rotation. Single guide rod and helical gear only can make driven helical gear \"10\" do unidirectional intermittent rotation. But driven helical gear \"10\" must be able to do continuously unidirectional rotation, so at least 3 or more guide rod and helical gear should be installed evenly on circumferential position along the axis of the driving shaft. Transmission Principle of the Swash Plate Pulse CVT.As shown in Fig.1, when the input shaft 1 rotating (or stationary), turn the speed screw 6, make the nut 5 move around, the fork of nut 5 drive slider 2 simultaneously move, the slider 2 drive connecting rod 4 through the hinge A, through the hinge A, the connecting rod 4 drive swash plate 3 rotate an angle around the input shaft 3. So that the input shaft rotates, the guide bar change the maximum displacement, the maximum displacement of guide bar is proportional to the output shaft speed, the greater displacement of guide bar, the higher the output shaft speed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003610_imece2013-63166-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003610_imece2013-63166-Figure2-1.png",
        "caption": "FIGURE 2: ANGLES",
        "texts": [
            " 1, a base coordinate frame, (OX ,Y,Z), is fixed at the center of the base platform with the Zaxis vertical and the X-axis pointing towards the point B1 in the horizontal configuration and towards the point C1 in the vertical one. In a similar way, the moving coordinate frame, (Pu,v,w), is defined with the w-axis normal to the platform plane and the uaxis pointing towards the point A1. According to this, the 3-DOF are a translation along the Z-axis and two rotations about the Xand Y- axes (\u03c8 and \u03b8 , respectively). Thus, the parasitic motions will be a rotation about the Z-axis (\u03c6 ) and two translations along the X- and Y-axes, (see Fig. 2). As the steps for deriving equations is similar for horizontal and vertical configuration, only the steps for the horizontal case are presented, and they are followed by the results for both the cases. The loop equation of the manipulator is given by Eq. (1): p = bi + li \u2212 rAi (1) Authors in [1] obtained the expression of the parasitic motions for the TZXY transformation matrix, but in this paper, TXYZ (case 1) and TYXZ (case 2) will be studied. For each case of Ttransf, the expressions of parasitic motions and the decomposition of the rotation in the mobile axes will be different"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.11-1.png",
        "caption": "Figure 12.11 A disc is falling by unwinding a ribbon.",
        "texts": [
            "228) Transforming this equation to the body coordinate frame, we have BF = BRG GF = m BRG GaB = m B GaB = m BaB + m B G\u03c9B \u00d7 BvB (12.229) We use the G-expression of the Newton equation (12.228) when the observer is in the G-frame. In this case, the observer is interested in the motion of the body with respect to herself. We use the B-expression of the Newton equation (12.229) when the observer is in the B-frame and moves with the body. In either case, the result of the analysis would be a set of differential equations. Example 733 A Wound Ribbon Figure 12.11 illustrates a ribbon of negligible weight and thickness that is wound tightly around a uniform massive disc of radius R and mass m. The ribbon is fastened to a rigid support, and the disc is released to roll down vertically. There are two forces acting on the disc during the motion, its weight mg and the tension of the ribbon T . The translational equation of motion of the disc is expressed easier in the global coordinate frame: \u2211 FY = \u2212mg + T = mY\u0308 (12.230) The rotational equation of motion is simpler if expressed in the body coordinate frame: \u2211 Mz = T R = BI B G\u03c9\u0307B + B G\u03c9B \u00d7 BIB G\u03c9B = I \u03b8\u0308 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002289_amr.591-593.569-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002289_amr.591-593.569-Figure3-1.png",
        "caption": "Fig. 3 Fig. 4 Fig. 5",
        "texts": [
            "0 ( rC : dynamic load rated of bearings) Z : the number of roller Finite element analysis on convexity for tapered roller bearings Selection of crown shape for tapered roller bearings. Convexity design of roller includes design and crown value calculation, but convexity shape research is the foundation. There are four kinds of generatrix shape in common use, including straight generatrix shape, arc generatrix shape, modification of arc generatrix shape and logarithmic profile[6]. Equivalent stress distribution of tapered roller bearings Fig. 3 with straight line, Fig. 4 with arc and Fig. 5 with arc-line modification profile Marginal effect occurred at the two ends of straight-line generatrix roller bearing, as revealed in Fig. 3. The equivalent stress between roller and inner circle was greater than outer circle, and the stress distribution was basically consistent with facts. The edge effect will be increased sharply when the load was unbalanced to accelerate fatigue damage of the roller bearing. In the Fig. 4, there was no edge effect on the two ends of arc roller bearing, of which the contact area was elliptic, and the contact stress distribution close to half an ellipsoid. But convexity value should be chosen reasonably, or the load will be unbalanced"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002764_0976-8580.99296-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002764_0976-8580.99296-Figure1-1.png",
        "caption": "Figure 1: RS - Model",
        "texts": [
            " Hussein [14] stated that an elastic\u2013plastic nonlinear analysis, under the dynamic load Adaptive Shifted Integration (ASI) technique, is capable of predicting the behavior of steel frame structures with reasonable accuracy. According to the literature review, contact analysis of a deformable sphere with a rigid flat using FE Analysis has been conducted by several researchers, and some of these studies consider the effect of material properties. The reserchers have also roughly taken the tangent modulus as 10% of the Young\u2019s modulus. Figure 1 shows that the RS-model (like the indentation approach). In the Brinell test a hard ball of diameter \u2018D\u2019 is pressed under a load \u2018W\u2019 into the plane surface under test. The present study aims to study the effect of strain hardness for single asperity contact parameters, for different materials, under the loading condition of the RS model. In the indentation process the indenter is a rigid member and the plate is a deformable member. In this study, the RS model chosen for analysis is based on the similarity of an indentation processes",
            " The loading relationship for the penetration depth is given by the relation \u03c9={9L2/8D}1/3[ 2{(1 \u2013 \u03bd2) / (E* + ET) }]2/3 (1) In Eq. (1), L is the applied load, D is the ball diameter, and the paired material constants \u03bd, E*, and ET are the Poisson\u2019s ratio, equivalent Young\u2019s modulus, and tangent modulus, respectively. E* is given by 1 1 11 2 1 2 2 2 / *E E E = \u2212 + \u2212  (2) In Eq. (2), 1 and 2 denote the material properties of the ball and plate, respectively. The projected surface diameter (d) of the residual impressed indentation is shown in Figure 1. The Table 1: Material properties E \u00d7 103 N/mm2 Y N/mm2 E/Y 120 69 1739.13 Journal of Engineering and Technology | Jul-Dec 2012 | Vol 2 | Issue 2100 following relationship gives the mathematical formula for calculating the diameter: d=2 [\u03c9 (D \u2013 \u03c9)] 1/2 (3) The material E/Y value of 552.63 is taken for observation of various parameters and it is related to the contact behavior of the sphere with flat (indentation approach), with incorporation of the tangent modulus. This is given in Table 3. Estimation of the contact area between the two consecutive steps of penetration is very important for analysis of the contact bodies in contact mechanics"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002330_20120619-3-ru-2024.00026-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002330_20120619-3-ru-2024.00026-Figure2-1.png",
        "caption": "Fig. 2. The BST-1 onboard Soviet Space Station Salyut-6",
        "texts": [
            " Matrosov had been carried in a large contribution on developing the basic research of dynamics and control for Soviet stratospheric and space astronomical observatories, Degtyarev (2007). So, together with A.S. Zemlyakov he had been applied the VLF method for research and designing the guidance and attitude control system by the Sun stratospheric (Krat and Kotlyar, 1976) observatory Saturn (Fig. 1, vertical size \u2248 70m), those by the Sky observatory Galaxy and by the space infra-red telescope BST-1 (Fig. 2, diameter of main mirror 1.5m) implemented onboard Soviet longtime orbital station Salyut-6. Developing mathematical models and research of such out-of-atmosphere astronomical observatories were supported by huge activity of V.N. Skimel, V.A. Strezhnev, A.M. Danilov, L.Yu. Anapolsky, G.G. Bilchenko, E.I. Druzhynin, V.M. Borodin, G.A. Kuzmin, A.I. Karpov et al. Obtained results were published in Danilov et al. (1975, 1976), Dul\u2019kin et al. (1983). Original project was also developed on designing a logic-digital control system for the Moon telescope guidance onto center of planet (the Earth) with incomplete phase, Bilchenko et al"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003002_amr.291-294.3282-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003002_amr.291-294.3282-Figure4-1.png",
        "caption": "Fig. 4 Normal and von Mises stresses with hardened depth of 25mm under 130t loads",
        "texts": [],
        "surrounding_texts": [
            "In order to compare the numerical results, the elastic Hertz theoretical results are also calculated, three different wheels with different hardening depths are applied, and three loads has been applied vertically to all cases. In order to simulate rail-wheel line-contact problems, crane wheel SYL800 and railway rail P50 have been chosen. The contour lines of normal, and von Mises stresses distribution of different wheels with hardening depth of 10mm, 25mm and whole hardened under 130t vertical load are shown Mises stresses of one contact point in wheel for different wheels with hardening depth of 10mm, 25mm and whole hardened along with the elastic Hertz theoretical results and the elastic numerical results under 100t vertical load. Table 2 and 3 give same information but under vertical loads of 130t and 150t, respectively. From Fig. 3, 4 and 5, it is shown that for line-contact case, the contact areas are in a shape of rectangular. This agrees with the Hertz\u2019s theory. From Table 1, 2 and 3, it is clear that the elastic numerical results for the contact stresses and contact area for line-contact case agree with the Hertz theoretical results very well. In line-contact cases, the normal stress field distribution and the von Mises stress field distribution in wheel under 130t vertical load are shown in Figure 3, 4 and 5. From these figures, the stress distributions for different hardening depth are very similar. depths at a point is listed along with the Hertz theoretical result and elastic numerical result under 100t vertical load. Compare to those of elastic wheel, the contact stresses decreased and the contact areas increased when the surface hardened wheel is used. Under 100t vertical load, the von Mises stress in the contact point is reduced by 14.4% together with 22.2% increasing in contact area when the wheel is surface hardened with a depth of 10mm. When the depth of surface hardening is increased to 25mm, compared with 10mm hardening depth wheel, the contact von Mises stress is decreased by 9.8% and the contact area is increased by 13.6%. When the wheel of whole hardened is used to simulate the contact problem, the contact von Mises stress is almost the same, and the contact area is exactly the same compared with the numerical results of wheel with hardening depth of 25mm. The similar things happen for the vertical loads of 130t and 150t according to the numerical results, which is listed in Table 2 and 3, respectively. support more loads because contact stresses are decreased by a large amount in the surface hardened wheel. For line-contact cases, the effect in stress decreasing of using surface hardened wheel is obvious. The increasing of hardening depth in the wheel will decrease the contact stress in a small amount further. When the hardening depth of the wheel reached to 25mm, the contact stresses will maintain a certain level though the hardening depth of the wheel increases. The contact stresses and contact areas will remain almost the same for a whole hardened wheel compared with the 25mm hardening depth wheel."
        ]
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.3-1.png",
        "caption": "Figure 4.3 Most frequent design of gear reducer: the axes of all shafts are in the same plane",
        "texts": [
            " Positions where the walls of a lesser mass are joined with those of a greater mass (bold part) should be preferably carried out with a gentle slope that rises in the direction of the increasing mass. Technologically, uniformity of cooling is ensured by a controlled cooling rate. Massive parts of housings, as well as works with poor thermal conduction, are cooled by means of metal inserts for heat dissipation. Reducers with cylindrical gears are usually designed in such a way that all the axes of the shaft lie in the final plane of the lower housing, which is connected with the corresponding areas of the cover (upper housing; hereinafter the assembly plane; Figure 4.3). Housings designed in this way are very suitable for mounting. Each gear shaft with bearings and all other elements that are contained in it can be assembled separately and independently of other shafts and then placed in the housing. To facilitate processing of the housing, the assembly plane is usually placed parallel to the base plane (Figure 4.3). However, there is another solution where the plane of assembly is not parallel to the base plane. This results in reduced weight and improved housing conditions of the gear lubrication, since larger gears of all steps are equally immersed in oil. For easier removal of the cover, the body is sometimes provided with threaded holes for depressing bolts. For the purpose of inspection, oil filling/draining and mounting of an all-level indicator or a thermometer, the corner and body are provided with drilled holes of the required form and size, shut with lids or plugs, nipples and so on"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001288_s11249-021-01479-x-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001288_s11249-021-01479-x-Figure5-1.png",
        "caption": "Fig. 5 Schematics of test specimens and sliding test setup",
        "texts": [
            " Figure\u00a04 presents the viscosity of the PAO8 oil plotted as Tribology Letters (2021) 69:111 111 Page 6 of 22 a function of temperature, as measured using a corn-rotor rheometer (AR2000, TA Instruments). The measurement shear rate was 100 1/s, the cone angle was 4\u00b0, and the measurement gap was 111\u00a0\u03bcm. Used PAO8 oil has a good compatibility to PA66 to ensure the long-term usage inside the automobile engine compartment at high temperature. Sliding tests were performed using the unreinforced PA66 or GF composite ring specimens and steel cylinders under PAO8 oil lubrication, according to the same method applied in the authors\u2019 previous works [27\u201331]. Figure\u00a05 presents schematics of the test specimens and tribometer setup. Four fixed steel cylinders were placed on the ring specimens, and a normal load was applied. Subsequently, the test specimens were rotated. Before the sliding tests, the sliding surface side of the ring specimens was dipped into the lubrication oil, and the oil lubrication was maintained during the sliding tests. A normal load was applied using the deadweights. When the sliding tests were conducted at 80 or 120\u00a0\u00b0C, a cartridge heater was inserted into the cylinder sample holder, and a constant temperature was maintained during the sliding tests"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003933_amr.328-330.520-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003933_amr.328-330.520-Figure3-1.png",
        "caption": "Fig. 3 Aircraft nose produced with SLM Fig. 4 Products formed by LENS 850 developed in MTT Corp.[12] by Optomec Corp. [12]",
        "texts": [
            " The same as SLS, SLM is only suitable for small parts with complex shape. Different from SLS, SLM can produce metal products with increased density significantly. To alleviate glomeration of molten metal, material parameter, process parameter and scanning mode should be controlled strictly. Hot isostatic pressing (HIP) is required to obtain high density parts, which often increases producing difficulty, time and cost. Part made from functionally gradient material(FGM) hardly is produced by SLS or SLM for its feeding way. Fig.3 shows the aircraft nose produced with SLM in MTT Corp [12] . Laser sintering method had been developed by using optical fiber laser, no subsequent heat-treatment after metal powder forming. Forming materials include powders such as high temperature alloy, stainless steel and die steel [12] . A 3D transient thermal finite model of selective laser melting process is developed to simulate distribution of temperature field in SLM. SLM method was developed for fabricating W-Ni-Fe alloys and 316L stainless steel parts"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.25-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.25-1.png",
        "caption": "Figure 10.25 Surface-mounted magnets and interior magnets: left, SPM motor; right, IPM motor. 1 \u2013 magnet; 2 \u2013 iron core; 3 \u2013 shaft; 4 \u2013 non-magnet material; 5 \u2013 non magnet material",
        "texts": [
            " A PM synchronous motor contains a rotor and a stator, with the stator similar to that of an induction motor, and the rotor contains the PMs. From the section on induction motors, we know that the three-phase winding, with three-phase symmetrical AC supply, will generate a rotating magnetic field. To generate a constant average torque, the rotor must follow the stator field and rotate at the same synchronous speed. This is also why these machines are called PM synchronous motors. There are different ways to place the magnets on the rotor, as shown in Figure 10.25. If the magnets are glued on the surface of the rotor, it is called a surfaced-mounted PM motor or SPM motor. If the magnets are inserted inside the rotor in the pre-cut slots, then it is called an interior permanent magnet motor or IPM motor. For a SPM motor, the rotor can be a solid piece of steel since the rotor iron core itself is not close to the air gap, hence the eddy current loss and hysteresis loss due to slot/tooth harmonics can be neglected. For the IPM motor, the rotor needs to be made out of laminated silicon steel since the tooth/slot harmonics will generate eddy current and hysteresis losses"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure7-1.png",
        "caption": "Figure 7. Combination Structure of steel part and plastics.",
        "texts": [
            " And the role of the suspension bushings is absorbed the forces caused ball joint. But because bushing outer pipe material is steel or aluminum, bushing and outer pipe occur to slip and each other is lack of pull-out strength. If it is covered plastics after nurring a steel pipe, steel pipes and plastics have a strong combination and pull-out strength degradation does not occur, as shown in Figure 6. If it is designed in shown a Figure 6, current forged upper arm of production is equal to pull-out strength. As shown in Figure 7, Ball joint pipe and bushing pipe were welded to center steel press and in order to increase in the durability and strength of welds, welds part were covered with plastics. Engineering plastics is materials that improved which thermal properties and mechanical strength. But even if it has a heat resistant, chemical resistant and high strength, engineering plastic is absolutely necessary considering of the safety factor properties because it cannot have permanent properties like steel and aluminum"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002797_amr.468-471.2141-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002797_amr.468-471.2141-Figure7-1.png",
        "caption": "Fig. 7 The bending of the cross-rail in the front-back direction",
        "texts": [],
        "surrounding_texts": [
            "Considering all the directions of the excitation, the simulated and experimental results and the error between the simulated and experimental results with joints have been provided in the Table.2 within the frequency of 100 Hz. It is found that the error between the simulated and experimental results of the whole assembled structure is less than 8.8%, while the error of the contact model often used in the existing literatures is one times bigger than the model.which indicates the proposed method in this paper is effictive. Table 2 Modal frequency and errors of the assembled structure Simulated result without joints [Hz] Simulated result with joints [Hz] Experimental result [Hz] Error [%] 23.318 17.689 19.388 -8. 763 70.602 50.418 46.409 8.638 90.211 84.851 81.439 4.190"
        ]
    },
    {
        "image_filename": "designv11_100_0002681_ijcnn.2012.6252849-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002681_ijcnn.2012.6252849-Figure10-1.png",
        "caption": "Figure 10. Growing topological Map Building",
        "texts": [
            " Neural gas has been also used for constructing a topological map, and furthermore, growing neural gas is used for incremental learning of the topological structure [18-19]. Local error measures are used for determining the place to insert new nodes. The competitive Hebbian rule generates the edges between nodes. In this paper, we propose topological mapping algorithm based on the concept of Growing Neural Gas (GNG) [20]. Mobile robot using measured distance data from LRF to obstacle for map building and localization (Fig. 9). Fig. 10 shows the procedure of topological map building. The first measure will be added as the initial location of measurement points on the topological map. Then, topological map information is updated by the measured data. When the point (node) in topological map is indicated as the reference vector by ri, Euclidean distance between input vector V and the i-th reference vector can be expressed as a equation (3). di = V ! ri (3) Next, the k-th output node to the minimum distance di is selected by the equation (4)",
            " In addition, i-th output node of reference vector is learned by the following equation, ri ! ri + \" #$ k ,i # (V % ri ) (5) where \u03be is the learning late (0<\u03be<1.0); \u03b6k,j is a neighborhood function (0<\u03b6k,j<1.0). The number of nodes, nnode is gradually increased when there is no node corresponding to input data within the threshold of distance. The number of inputs in each sampling of the distance information is L (L=682), and this method is composed of three steps of node addition, learning, and node deletion. Fig. 9 and Fig. 10 show an example of creating a topological map. We show the procedure of the topological map building; The node deletion is performed in order to remove unnecessary and crowed nodes (TABLE II). Fig.10 shows an example of topological map building. If the node does not exist in the position corresponding to the measured distance, a node is added to the map (Fig. 10 (a)). If there are many nodes crowded, some of them are removed from the map (Fig. 10 (b) and Fig. 10 (c)). In this part, we think about SLAM in living space. When a person passes through the measurement area of the robot, LRF sensor cannot be separated whether person or static object. The person information will remain as traces of noise on the map. For example, Fig. 11 is an example when a person has passed the measurement range. In Fig. 11 (c), dynamic object leave one\u2019s trace as a noise on the map. In order to solve such problems, noise reduction is performed by the following equation. di = {li (t + j)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure11-1.png",
        "caption": "Figure 11. 3-Side Gate Mold Flow Analysis.",
        "texts": [
            " As shown in Figure 10 and Table 1, composite material strength increased to 15 percent compared to forged steel and stiffness of important factor of suspension performance increased to 5 percent compared to forged steel. Because there should be evenly distributed injection glass fiber in order to insert engineering plastics, mold flow analysis must be accompanied. Engineering plastic injection gate is 3-Side, injection pressure is approximately 40MPa, injection time is 1.7sec, plastic melting temperature is approximately 280 degree and mold temperature is approximately 80 degrees was developed by setting and mold as shown in Figure 11. In the upper arm CAE analysis, the result of composite material stiffness is more than the equivalent level compared to forged steel. Front suspension module CAE modeling applying composite material suspension is made as shown Figure 12. So lateral force happened to vehicle apply to CAE modeling and lateral stiffness analysis of front suspension module has been implemented. The result of composite material suspension module of lateral stiffness is more than the equivalent level compared to Forged steel suspension module as shown in Figure 13"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003419_amr.199-200.449-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003419_amr.199-200.449-Figure1-1.png",
        "caption": "Fig. 1 Measure system of the rotational rectangular coordinates",
        "texts": [
            " Unlike the fore-mentioned optical measure method, a new method is proposed by Hirano. In this method, a magnetized ball is assembled into a bearing, and then the voltage of the search coil caused by the variation of the magnetized ball\u2019s magnetic axis is measured. In the 80s, Kawakita [4-11]et al. developed Hirano\u2019s method. A 3D measure method which utilizes magnetic sensitive semiconductor elements, namely Hall elements, as the sensors to measure rolling elements\u2019 motion is proposed. A magnetic circuit (refer to fig. 1) is composed by the magnetized balls and the Hall elements which are buried in the cage and the outer race. The variation of the magnetized balls\u2019 magnetic axis is measured and used to derive the balls\u2019 3D motion. This method (rectangular coordinates method) solves some problems existed in the former methods and was considered as the only proper method to obtain the 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 TTP, www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003946_amm.120.343-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003946_amm.120.343-Figure2-1.png",
        "caption": "Fig. 2 Force diagram of the inner frame under the radial load",
        "texts": [
            " (2) where wr is the number of wire bearing the radial loads from one row of rolling elements. It is defined by wr=2 for journal and thrust bearings. C is the load distribution factor of a wire ball bearing under radial loading. It is defined by C=4 for line contact bearings with zero gap, C=4.37 for point contact bearings with zero gap, C=5 for bearings with proper clearance. In general, the load distribution factor of the wire race ball bearing is equal to C=5. Next, we can perform the mechanical analysis of the ball labeled B1 which has the maximum load. Fig. 2 shows the force diagram of the inner frame under the radial load. According to the overall and local mechanical analysis of the contact between the ball labeled B1 and the wires, the following equilibrium equation can be established 3 3 4 4sin sinF F\u03b1 \u03b1= . (3) 1 1 2 2sin sinF F\u03b1 \u03b1= . (4) 1 1 2 2 3 3 4 4cos cos cos cosF F F F\u03b1 \u03b1 \u03b1 \u03b1+ = + . (5) 1 1 2 2 1cos cos BF F F\u03b1 \u03b1+ = . (6) To solve the radial stiffness of the wire race ball bearing, the relationship between the equivalent springs should be clear"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001797_qr2mse.2013.6625695-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001797_qr2mse.2013.6625695-Figure1-1.png",
        "caption": "Figure 1. The lock mechanism Figure 2. The lock mechanism in close position. in open position.",
        "texts": [
            " The test system and test cramp designed for the test method can load the combined environmental stress and working load, and can simulate the close-open function of the lock. The reliability assessment method for single sample is also proposed. At last, with the example of landing gear door lock of an airplane, the using of this method is explained. The method proposed in this paper can also be used for other similar mechanical product. II. THE FUNCTIONAL ANALYSIS OF THE LOCK The lock studied in this paper is composed of lock body, and rocker arm, and piston, and piston rod, and rod-1, and rod2, and latch hook, which is shown in Fig.1. In the open process, under the pressure of the hydraulic oil, the piston and its rod will move toward right, and the rocker arm will turn anticlockwise. At the same time, the latch hook will also turn anticlockwise until the lock is closed, under the drive of rod-2 and rocker arm. Similarly, in the close process, the piston will move toward left and the latch hook will turn clockwise until the lock is open. The lock in close position is shown in Fig. 1, and the open position is shown in Fig. 2. A typical motion cycle of the lock is composed of close the lock (in take off process), and keep the close position (in flying process), and open the lock (in landing process). 978-1-4799-1014-4/13/$31.00 \u00a92013 IEEE 817 The lock can be repaired in principle, but for the gear door lock is a key component of the landing gear system, serious accident may be caused once there is a failure in the lock. As a result, the lock is defined as non-repairable product, and the overhaul period of the airplane is defined as the design life of the lock"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003427_epe.2013.6634733-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003427_epe.2013.6634733-Figure7-1.png",
        "caption": "Fig. 7: Lithium Battery Charger - 48V/50A [6]",
        "texts": [
            " The balancing MOSFET is screwed on a surface of the box and connected by wire therefore it is not placed in PCB. A photo of the Monitoring Unit in the aluminium box and the placement of it in the kart are shown in Fig. 6. When the charger is connected the BMS unit controls the charging. The lithium battery charger POW48V50AT made by GWL Power Company [6] is used for charging. The nominal output current of this charger is 50 A. It allows achieving the full charge of all batteries in approximately 1.5 hour. The charger is shown in the Fig. 7. The battery state of charge (SoC) is calculated by this BMS unit. It is calculated by the integration of the measured battery current. The BMS unit also allows logging of selected operational values (temperatures, currents, voltage, RPM, traveled distance, velocity and battery state of charge) to an SD card for further processing. It is possible to connect PC by USB through this unit and download the logged data. The BMS unit transmits CAN message which includes these values: battery current, 4 battery voltages, 4 battery temperatures, SoC and error flags"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003665_isie.2013.6563839-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003665_isie.2013.6563839-Figure3-1.png",
        "caption": "Fig. 3. Relation Between Camera and Image plane",
        "texts": [
            " Image Processing The stereo camera are mounted on the robotic wheelchair, and these image are used for robot motion estimation. Each image is divided to small cell blocks as shown in Fig.2, and DV between left and right images, and OFV in consecutive left images are calculated respectively by correlation matching algorithm. The horizontal disparity di yields the z-axis position of corresponding block area by (1). Zi = L di f (1) where L and f are the interval of stereo cameras and focal length respectively. Subscript i means the block number. In Fig.3, the camera coordinate and several parameters are specified. Assuming that the target viewing point Pi = (Xi,Yi,Zi) is stationary in world coordinate, the relation between camera motion and Pi in camera frame can be written as (2). P\u0307i = dPi dt = \u2212(V + \u03a9\u00d7Pi) (2) where V and \u03a9 are the translational camera velocity [Vx,Vy,Vz]T and rotational velocity [\u03c9x,\u03c9y,\u03c9z]T respectively. When the point Pi is projected on the image plane at pi = (xi,yi), the relation between Pi and pi is given by (3). xi = f Xi Zi , yi = fYi Zi (3) The time derivative of (3) can be calculated as follows: ui = dxi dt = 1 Zi ( f dXi dt \u2212 xi dZi dt ) (4) vi = dyi dt = 1 Zi ( f dYi dt \u2212 yi dZi dt ) (5) The vector ui = [ui,vi]T is an optical flow vector at pi, by substituting (2) into (4) and (5), the final formulation is obtained as (6)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.39-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.39-1.png",
        "caption": "Fig. 2.39 4PaCP-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology Pa\\C\\P",
        "texts": [
            "21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001202_tec.2021.3058804-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001202_tec.2021.3058804-Figure2-1.png",
        "caption": "Fig. 2. Induction machine geometry: (a) cross-sectional view over one polepair, (b) winding configuration.",
        "texts": [
            " An illustration of a reconstructed \u03bb(t) is provided in Fig. 1(c). The elements of the time-averaged inductance matrices and the flux ripple vector are obtained using (28). In particular, the flux ripple vector \u03bb\u0303\u2212 qdr0 can be expressed as in (10) with k representing a tuple (u, v), and the elements of |\u03bb\u2212 k | and \u03c6\u2212 \u03bb,k corresponding to 2|\u2113u,v|/XY and \u03c6u,v , respectively. The proposed model is validated using a three-phase 4-pole (p = 2) squirrel-cage IM as a case study. The cross-section of the machine geometry over one pole-pair is shown in Fig. 2. The machine has 48 stator slots and 70 rotor slots, i.e., the number of rotor bars per pole-pair is n = 35. The rotor cage is made of copper. The end-ring segment resistance and inductance are Re = 0.19 \u00b5\u2126 and Le = 4 nH, respectively. A nonlinear M250-35A steel grade [8] is used for the stator and rotor core materials. A lamination stacking factor of 100% is Authorized licensed use limited to: London School of Economics & Political Science. Downloaded on May 16,2021 at 23:17:07 UTC from IEEE Xplore"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003946_amm.120.343-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003946_amm.120.343-Figure3-1.png",
        "caption": "Fig. 3 Equivalent model of the contact between a ball and the four wire races",
        "texts": [
            " According to the overall and local mechanical analysis of the contact between the ball labeled B1 and the wires, the following equilibrium equation can be established 3 3 4 4sin sinF F\u03b1 \u03b1= . (3) 1 1 2 2sin sinF F\u03b1 \u03b1= . (4) 1 1 2 2 3 3 4 4cos cos cos cosF F F F\u03b1 \u03b1 \u03b1 \u03b1+ = + . (5) 1 1 2 2 1cos cos BF F F\u03b1 \u03b1+ = . (6) To solve the radial stiffness of the wire race ball bearing, the relationship between the equivalent springs should be clear. According to the equivalent model of the contact between a ball and the four wire races in our previous work [1], (see Fig. 3), the total equivalent spring is constructed by two sets of equivalent springs in series. One set consists of the No. 1 and No. 2 equivalent springs which are in parallel at an angle, the other one is made of the No. 3 and No. 4 equivalent springs are in parallel at an angle. The deducing process of the above stiffness is similar with the axial equivalent stiffness in our previous work [1]. Hence, the radial equivalent stiffness expression of the ball having the maximum load can be directly presented as follows 12 34 1 12 34 r K K K K K = + "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure5-1.png",
        "caption": "Figure 5. FE model of chassis assembly with Super element",
        "texts": [],
        "surrounding_texts": [
            "Other than the identified critical regions are modeled with super elements using sub-modeling technique available in MSC / Nastran. These super elements are reduced to stiffness and load vector to interface DOF of these critical regions. Similar linear static analysis is carried out for identified critical regions along with reduced stiffness and load vector and compared the results with and without super elements. Boundary conditions for the identified critical regions are captured from super element without any user intervention."
        ]
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.46-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.46-1.png",
        "caption": "Figure 12.46 A uniform rotating arm with a hanging weight m2 at the tip point.",
        "texts": [
            " To solve the dynamics of the four-bar mechanism, we must calculate the accelerations 0ai and 0\u03b1i and then find the required driving moment 0M0 and the joint forces. Example 760 A Turning Arm with a Tip Mass Carrying masses are regular functions of mechanical machinery. The carrying mass will change the position of the mass center as well as the mass moment properties of the machine. To see the effect of such a massive load on the required actuating force systems, let us consider the uniform arm of Figure 12.46(a) with a hanging mass m2 at the tip point. Figure 12.46(b) illustrates the FBD of the arm. Adding m2 to the system moves the mass center of the arm to 1r1: 1r1 = m1 m1 + m2 \u23a1 \u23a2 \u23a3 l/2 0 0 \u23a4 \u23a5 \u23a6+ m2 m1 + m2 \u23a1 \u23a3 l 0 0 \u23a4 \u23a6 = \u23a1 \u23a2\u23a2 \u23a3 m1 + 2m2 2 (m1 + m2) l 0 0 \u23a4 \u23a5\u23a5 \u23a6 = \u23a1 \u23a2 \u23a3 rx 0 0 \u23a4 \u23a5 \u23a6 (12.691) The relative position vectors m and n of arm (1), which is the only link of the system, are 1n1 = \u2212 1r1 = \u2212rx \u0131\u0302 (12.692) 1m1 = l\u0131\u0302 \u2212 1r1 = (l \u2212 rx) \u0131\u0302 (12.693) 0d1 = \u2212 1n1 + 1m1 = l\u0131\u0302 (12.694) 0m = 0R1 1m1 = \u23a1 \u23a3 (l \u2212 rx) cos \u03b8 (l \u2212 rx) sin \u03b8 0 \u23a4 \u23a6 (12.695) 0n = 0R1 1n1 = \u23a1 \u23a3 \u2212rx cos \u03b8 \u2212rx sin \u03b8 0 \u23a4 \u23a6 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003534_amr.308-310.2220-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003534_amr.308-310.2220-Figure1-1.png",
        "caption": "Fig. 1 Cycloid gear and output mechanism assembly model",
        "texts": [
            " In order to verify the contact state between FA cycloid drive transmission output dowel pin and cycloid gear pin hole, taking the FA45-29 model for example, this article fulfilled a special position\u2019s three-dimensional contact finite element analysis in the whole course of transmission (a the dowel pin exists at \u03b1T=90\u00b0position, \u03b1T as the first contact with the dowel pin Angle, this position is corresponding to a needle tooth existed position [3] ). First using three-dimensional modeling software POR/E establish FA cycloid drive transmission whole assembly model ,and then using the seamless connection interfaces between POR/E and ANSYS, put the built assembly model to the ANSYS, then obtained the geometrical model shown as in the Fig.1. Finite element grid partition Considering various factors may influence on the quality of the grid partition and the geometrical characteristic of cycloid gear and output dowel pin, using the body sweeping grid partition method, and manual control of the grid partition, the finite element mesh model is finished shown as in the Fig. 2. Among them, part a) is the whole finite element model, part b) is the local finite element model, node number is 455756, unit number is 437612. All rights reserved"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure2-1.png",
        "caption": "Figure 2. Stresses on truck chassis for vertical-3Gload case",
        "texts": [
            " Aggregate masses, load body and payload masses are modeled as lumped mass at respective CG locations and are connected to the structure using RBE2 and RBE3 elements. The pin joints in the suspension are simulated using CBAR and RBAR elements to allow rotational degrees of freedom. FE model of frame assembly is shown in Figure 1. The wheel points are constrained such that the structure is subjected to minimum constraints. Linear static analysis is carried out to evaluate the stresses on FSM for vertical 3g loading condition. FE model is solved using MSC / Nastran and results are post-processed using Altair / Hyper view. Figure 2 shows the stresses on the chassis assembly. The below mentioned regions are identified as critical regions where linear assumptions are invalid because load transfer happens through surface contact interactions. \u2022 Rear Spring Front Bracket (RSFB) \u2022 Rear Spring Middle Bracket (RSMB) \u2022 Rear Spring Rear Bracket (RSRB) Hence these identified critical regions are considered for submodeling. SUB-MODELING TECHNIQUE [2] Consider a finite element model of a structure that is subdivided in two parts or substructures A and B"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002643_amr.308-310.1714-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002643_amr.308-310.1714-Figure1-1.png",
        "caption": "Fig. 1 3D drawing of the transmission device",
        "texts": [
            " However, miniature shaft couplings and holders now available are mostly with complex structure, large size, high inertia moment, poor compensation capacity of misalignment of the shafts and difficult to install, and could not meet the requirements of increasingly developing miniature high-speed precision machinery and instruments. Aiming at above shortages, a new transmission device was innovatively presented and its design calculating models were founded. The transmission device was applied to an ultra-high speed micro-spindle to assess and verify its performance. Fig. 1 shows the three dimensional (3D) drawing of the transmission device, which consists of an elastic shaft coupling integrated with the drive shaft, a collet with 4 jaws in the end of the drive shaft and a clamp ring. The driven shaft is installed in the collet and then the clamp ring placed around the outside of the collet. The locations of the driven shaft and the clamp ring are determined by step surfaces. The elastic shaft coupling is manufactured by cutting opposite grooves in the drive shaft, which forms thin walls"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002655_amr.789.443-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002655_amr.789.443-Figure2-1.png",
        "caption": "Figure 2: Schematic Diagram of Equipments Figure 3: LCS with 21 Lamps",
        "texts": [
            " On the one of the end tip of the shaft, a 24 cm arm was placed horisontally to measure torque produced by the generator. A weight scale was placed on the other end tip of the torque arm to measure force of the end tip of the torque arm. Driving System. The driving system drove the generator by belt and pully transmission. In addition, the driving system also could adjust the rotation speed from 0 up to 450 rpm. Main power supply used 1 phase electric source. The scheme of experiment equipments was described in Fig. 2. The voltage regulator was used to stabilize voltage on 220 volt during the experiment. The specification of voltage regulator used in the experiment was KENIKA, Automatic AC Voltage Regulator AR-1200. A contoller was used to adjust the rotation speed of the driving motor. The specification of controller used in the experiment was TOSHIBA transistor inverter, VFS112055PM-WN (R5), 5.5 kW with 1 phase input and 3 phase output. The controller adjusts the speed of electric motor from 0 up to 1600 rpm"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002199_978-1-4419-9323-6_4-Figure4.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002199_978-1-4419-9323-6_4-Figure4.16-1.png",
        "caption": "Fig. 4.16 Example of turbine control",
        "texts": [
            " The turbine or the rotor speed is continuously adjusted as the wind velocity changes in order to keep Cp at the optimal value, and the output power at its maximum value. The maximum power tracking characteristic like the one shown in Fig. 4.5 for a given turbine design can be obtained from the wind power model, and can be used in the design of turbine control. Following the maximum power tracking characteristic, for a given wind speed, the maximum power and the rotor speed are fixed. Hence, one of the control methods is to give the power command as a function of the rotor speed. For example, consider the operating point A in Fig. 4.16 at a wind speed of 10 m/s. If the wind speed increases to 12 m/s, the power from wind increases as indicated at point B while the electric power output has not yet increased. This accelerates the rotor increasing the rotor speed, and therefore, the electric power command. The operating point moves along the characteristic corresponding to wind speed of 12 m/s as shown in the figure till the point C is reached, where the electric power has increased to the maximum possible value for wind speed of 12 m/s"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002278_2013-01-1491-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002278_2013-01-1491-Figure8-1.png",
        "caption": "Figure 8. Run out at Pitch Circle Diameter of Dog Teeth",
        "texts": [
            " The maximum tilting of the 2nd gear with maximum radial gap between the gear bore diameter and outer diameter of the needle roller bearing is calculated from the dimension list given in Table 1. The maximum angular tilt of the 2nd gear depends upon the minimum effective supporting length of needle roller (b_min) and the minimum package diameter of rollers (d_min). The b_min and d_min are calculated from the given dimensions and then calculate the gear tilting angle \u20180.268\u00b0\u2019 (4.682 mill rad) as shown in Figure 7. Run out due to the tilting angle \u2018z\u00b0\u2019 of the 2nd gear is calculated from the given dimension in Table 1. Figure 8 shows the mounting position of the 2nd gear in gearbox and the maximum tilting position of the 2nd gear due to the radial gap between the 2nd gear bore diameter and outer diameter of the needle roller bearing. The initial position of the dog teeth at the pitch circle diameter \u2018P009\u2019 from the center point \u2018C\u2019 of the gear barrel width is AB and at an angle of \u2018u\u00b0\u2019. After the maximum tilting \u2018z\u00b0\u2019 of the gear due to the maximum radial gap, position of the dog teeth at pitch circle becomes A\u2032B\u2032 from the center point \u2018C\u2019"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003876_aqtr.2012.6237754-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003876_aqtr.2012.6237754-Figure1-1.png",
        "caption": "Figure 1. PARASURG 9M structure",
        "texts": [
            " In such applications where the robot interacts directly with the human (both surgeon and patient), safety is the most important aspect. In [5] it was shown that over 60% of the requirements of a surgical robot refer to safety issues. The same rule applies to the control system of such a robot: to be robust and most important, to prevent any incident that may hurt the persons surrounding it. Paper [7] presents an interesting algorithm for implementing the reverse engineering concept for medical applications. The parallel hybrid robot PARASURG 9M structure is presented in Fig.1. The robot consists of two subsystems: PARASURG 5M [6] and the active robotic surgical tool: PARASIM. PARASURG 9M system was designed in such a way that the connection between the two subsystems ensures the data transfer compatibility and feasibility. The two units share the data in a server client transfer mode under the supervision of the central control system. II. HARDWARE STRUCTURE PARASURG 9M design uses a parallel structure with 5 motors (PARASURG 5M), presented in Fig.2. It can control both an endoscope (laparoscopic camera) or an active surgical instrument (in this case PARASIM). PARASURG 5M has 5 actuators, three of them are installed at the base platform of the robot and the other two integrated in the orientation module structure (Fig.1). The structure of PARASURG 5M is build in such a way that PARASIM can be easily interconnected [8]. PARASIM robotic instrument model was developed using CAD, integrating a spherical guiding mechanism (Fig.2) The system is actuated by a series of DC motors, and motion transmission elements (belts, gears) in order to transmit the 3+1 degrees of freedom: q6, q7 and q8 (rotations) as well as one translation \u2013 q9 (acting as a griper). Fig.3 shows the position of the actuators which control the end effector"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003161_amm.105-107.244-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003161_amm.105-107.244-Figure3-1.png",
        "caption": "Figure 3 Scheme of used roller",
        "texts": [
            " The device was loaded by the mass element located in the middle of both tables \u2013 symmetrical loading of 1000kg was assumed. The model contained 12726 finite elements and 33662 nodes. The two types of materials were assumed in the model \u2013 construction steel E=210000MPa. \u00b5=0.3 and wood E=10000MPa. \u00b5=0.2 which is located in the form of wooden board on the top sides of both tables. The fundamental part of the solution is the realization of nonlinear joint between the tables and guiding U-profiles. The physical realization of connection of tables and guiding profiles is obvious from Fig. 1 and Fig. 3. The tables are provided on both sides by two rollers (Fig. 3) guided in Uprofiles. So the longitudinal and lateral guidance is ensured. The lateral vibration is avoided due to the side areas of U-profile. In case of longitudinal vibration only the compression force can be transferred because the contact between the roller and U-profile occurs only at one its side. This fact is to be considered in FE model. The transfer of the forces between roller and guiding profile was realized by coupling of degrees of freedom belonging to both parts. The rise of compression force in lateral direction is ensured in all cases (at one or other parallel side of U-profile)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002416_icma.2013.6618014-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002416_icma.2013.6618014-Figure2-1.png",
        "caption": "Fig. 2 Microphone array with 5 microphones L - Q.",
        "texts": [
            " A technique for the real-time detection and the identification of a particular voice among plural sounds using a microphone array, based on the location of a sound source and its tonal characteristics will be described. The algorithm is then installed in a robotic arm, which tracks a particular sound source in the midst of plural sounds, and interacts with a particular human speaker. The authors are working for the development of a robotic auditory system as shown in Fig. 1, which consists of a microphone array, low-pass filters (LPFs), a computer with a A/D card, and a robotic arm. The microphone array composed of 5 microphones L - Q arranged diagonally as shown in Fig. 2 is settled at the tip of the robotic arm, so that the robot actively moves in a 3D space for listening to and interacting with sounds around it. The microphone array is connected to the computer via the A/D card to input sound signals. The sampling frequency was set to 16.0 kHz, and the analysis window of 1024 points was chosen for the analysis of acoustic characteristics of inputted sounds in this study. The cut-off frequency of the LPF, which was placed in the inlet of the A/D card, was set to 8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002153_cbo9780511760723.019-Figure17.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002153_cbo9780511760723.019-Figure17.9-1.png",
        "caption": "Fig. 17.9 (a) Saturation in a magnetically soft material with no hysteresis. This is representative of superparamagnetism, the property magnetic beads are designed to exhibit. Superparamagnetism is typically observed in small domains of ferromagnetic materials. (b) Saturation in a magnetically hard material with hysteresis, such as iron or magnetite at long length scales. (After [61].)",
        "texts": [
            " Columbia University Libraries, on 27 Nov 2018 at 09:12:43, subject to the Cambridge Core terms of use, available at Saturation is a key property that prevents magnetophoresis from being directly analogous to DEP. Materials have a saturation magnetism (e.g., 2.2T for iron) that corresponds to their spins being fully aligned. Increased magnetic field does not lead to further spin or further magnetic flux density.4 This leads to a magnetic permeability that is dependent on the applied magnetic field (See Fig. 17.9a). In magnetically soft materials, the B-H curve5 is nonlinear but shows no hysteresis. In magnetically hard materials, the B-H curve also shows hysteresis, as seen in Fig. 17.9b, owing to the fact that nonequilibrium states are metastable. 17.2.3 Magnetic properties of superparamagnetic beads Superparamagnetic beads typically consist of a polystyrene matrix (of roughly 1\u00b5m diameter) filled to about 15% mass fraction with roughly 10nm ferromagnetic or ferrimagnetic particles. The 10-nm particles are described as superparamagnetic, which means that they respond strongly to external magnetic fields but do not retain permanent magnetism and do not show hysteresis. Owing to their small domain size, the nanoparticles can orient and equilibrate themselves quickly, so they show strong effects like a ferromagnetic or ferrimagnetic material and avoid hysteresis and permanent magnetism"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003770_msf.706-709.2950-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003770_msf.706-709.2950-Figure1-1.png",
        "caption": "Fig. 1. (a) The TTT experimental arrangement and (b) the positions of the burners relative to the weld centre line and welding torch (\u201cA\u201d refers to the leading case, \u201cB\u201d to the parallel situation and \u201cC\u201d to the trailing case).",
        "texts": [
            " The burners have eight nozzles with an interdistance of 30 mm and operate with acetylene and compressed air. In this study, the pressure of acetylene and compressed air was set to 0.5 bar and 1.5 bar respectively. The burners are calibrated for different maximum temperatures. After calibration of the burners, TTT during welding was performed with different positions of the burners with respect to the weld centre line and the welding torch. The burners were positioned parallel to the weld centre line and moved with the same speed as the welding torch. Fig. 1a shows the TTT experimental arrangement, while the locations of the burners relative to the weld centre line and the welding torch are shown in Fig. 1b. The out-of-plane deformation of the plates both for conventional and TTT welding was measured to find the optimum conditions using a Digital Image Correlation method (DIC). The temperature and the residual stresses are measured for this optimized case. The temperature of the plate is measured using glass insulated k-type thermocouples to an accuracy of \u00b12.2 oC attached at the rear surface along a line perpendicular to the welding direction. Residual stress profiles of the conventionally welded plates and the plates after welding with side heating were measured at the Paul Scherrer Institute (PSI) by neutron diffraction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure3-1.png",
        "caption": "Figure 3 \u2212 Wheel multibody submodel",
        "texts": [
            " The headstock is constrained to the ground by means of a dummy body that allows the castor to translate; the vertical and the lateral translations are counteracted by the castor weight and by an horizontal spring, respectively. A second dummy body is connected to the ground by means of a linear actuator to impose an advance motion, with an assigned forward constant velocity. The geometric and inertial body properties were experimentally evaluated and are reported in table I. The tire-road interaction has been modelled by the Pacejka formula [6]. To determine the tire point where is applied the interaction force the \u201ctwo link multibody sub-model\u201d (Fig. 3) described in [7], was use. It allows to geometrically pick out the position of the kinematic tire-ground contact point taking into account the actual geometry of the tire transverse profile. The sub-model is constituted by two links joined by means of an hinge in point C (Fig. 3). The lower link is always perpendicular to the ground and its free end is located in the theoretical contact point. The upper link is connected to the wheel axis and can follow the wheel camber rotation. If the tire profile is different from a circular torus one, the lower link length varies with the camber angle (remaining superimposed 2 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/75795/ on 04/09/2017 Terms of Use: http://www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002275_amm.371.250-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002275_amm.371.250-Figure4-1.png",
        "caption": "Fig. 4. Internal wax pressure simulation, using SolidWorks Simulation soft program: a) Fixtures on cylindrical faces; b) Surface fixtures",
        "texts": [
            " 1 and after the \u201cCATALYST\u201d FDM system software simulations, the authors made, step by step, various design transformations concerning the industrial parts \u201dHalf-mold Lever\u201d connected with the specific capabilities of the RP technologies. There have been several attempts on simulations to determine suitable part deposition orientation for different objectives like dimensional accuracy, build time, support structure and his effect, etc. (Table 1 and Fig. 2). pressure some simulation tests using SolidWorks Simulation software were done for different materials and for different RP technologies (Table 2, Fig. 3 and Fig. 4). for this sort of parts are CT 4=0.36mm, CT5=0.5mm, CT6=0.7mm. Prototyping machines (naturally and the technologies too) the authors has built a number of \u201cHalfmold Lever\u201d at University of Applied Science of Aachen, Germany laboratories area. The used RP technologies was as follows; FDM-Fused Deposition Modeling, \u201dDimension sst\u201d machine; 3Dprintig, \u201dZ printer 450\u201d machine and PolyJet, Objet 30 machine. The half-molds, the parts, the materials and theirs dimensions, built in this RP ways, may be seen in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003210_itsc.2011.6082926-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003210_itsc.2011.6082926-Figure1-1.png",
        "caption": "Fig. 1. Single track model with roll degree of freedom [1]",
        "texts": [],
        "surrounding_texts": [
            "I. INTRODUCTION\nNoting its economical cost and serious impact on human safety, rollover prevention has nowadays become an important area of research [1]\u2013[8]. In order to predict rollover situations, previous studies employ several methods, some of which include the use of the load transfer ratio [1], [2], [3] the time to rollover metric [4], and lateral acceleration thresholds [5]. Differential braking [1], active steering [6], active suspension [8], and active roll stabilizer bars have been used as actuators to prevent rollover. The roll angle, roll rate, slip angle, lateral acceleration, center of gravity height have all been considered to be very important parameters that are highly correlated with vehicle rollover.\nThe focus of this paper is vehicle parameter identification which has individually drawn a lot of research interest. In the MMST (Multi Model Switching and Tuning) methodology [1], [11], multiple vehicle models with different parameter values (CG height and other unknown parameters) are considered. The parameters of the model which is closest to the real vehicle are taken as the estimated parameters which include the height of center of gravity. In another study, extended Kalman Filter, extended Luenberger observer and sliding mode observer were used to estimate vehicle side slip angle and velocity [9]. Although vehicle velocity and side slip angle are estimated in [9], the focus of this paper is to identify the CG height and several other unknown parameters of the vehicle based on Kalman filtering techniques that can effectively estimate the state of a linear system. To this end, the parameter identification problem is first converted to a state observation problem for a discrete-time, time-varying system with yaw rate and lateral acceleration as the outputs. As in [1], the recursive least squares techniques can also be used in parameter identification. However this option is not viable for the problem under consideration in this paper as the state vector of the system is speed dependent\nThis work was supported in part by TUBA GEBIP programme and Bogazici University Research Fund.\nThe authors are with the Department of Electrical and Electronics Engineering, Bogazici University, Istanbul 34342, Turkey. mehmet.akar@boun.edu.tr\ntherefore time-varying; e.g., when we couple the rollover controller with parameter identification. On the other hand, the proposed method can estimate the relevant parameters effectively in such cases.\nThe organization of the paper is as follows. In Section 2, the mathematical vehicle model is presented. In Section 3, the vehicle parameter identification problem and the proposed method for parameter identification are discussed. The efficacy of the proposed algorithms is evaluated via simulation results in Section 4. Finally, some concluding remarks are given in Section 5.\nII. VEHICLE MODEL AND PARAMETERS OF INTEREST\nThe mathematical model of the system to be controlled is\ngiven by [1]\nx\u0307 = Ax + B\u03b4\u03b4 + Buu (1)\nwhere\nA =\n\n   \n\u2212\n\u03c3Jxeq mJxx\u03c5\n\u03c1Jxeq\nmJxx\u03c52 \u2212 1 \u2212 hc Jxx\u03c5 h(mgh\u2212k) Jxx\u03c5\n\u03c1 Jzz\n\u2212 \u03ba Jzz\u03c5\n0 0\n\u2212 h\u03c3 Jxx\nh\u03c1 Jxx\u03c5\n\u2212 c Jxx\nmgh\u2212k Jxx\n0 0 1 0\n\n   \n(2)\nB\u03b4 = [ C\u03c5Jxeq\nmJxx\u03c5 C\u03c5l\u03c5 Jzz \u2212 hC\u03c5 Jxx\n0 ]T\n(3)\nBu = [ 0 \u2212 T 2Jzz 0 0 ]T\n(4)\n978-1-4577-2197-7/11/$26.00 \u00a92011 IEEE 1440",
            "and the state x contains side-slip angle \u03b2, yaw rate \u03c8\u0307 , roll rate \u03c6\u0307 and roll angle \u03c6, i.e.,\nx = [ \u03b2 \u03c8\u0307 \u03c6\u0307 \u03c6 ]T\n(5)\nIn order to simplify the system equation, auxiliary parameters \u03c3 = C\u03c5 + Ch , \u03c1 = Chlh \u2212 C\u03c5l\u03c5 , \u03ba = C\u03c5l2\u03c5 + Chl2h, and the equivalent roll moment of inertia Jxeq = Jxx + mh2 are defined. The other parameters used in the above equations are tabulated in Table I.\nThe steering input \u03b4 directly affects both roll and lateral dynamics of the vehicle as expected. The second input u is the total differential braking force on the wheels which will be the controller\u2019s input to the system. According to the Newton\u2019s law, the braking force will change the speed of the vehicle. The differential braking governs the vehicle longitudinal speed as \u03c5\u0307 = Fx\nm \u2212 |u| m where Fx is the\naccelerating force in the longitudinal direction.\nA. Vehicle Parameters\nThe vehicle parameters in (1) can be categorized in two groups: the lateral dynamics parameters C\u03c5 , Ch, l\u03c5 , and the roll dynamics parameters k, c, h. In this paper, Kalman filtering and recursive least squares techniques are used to estimate these parameters.\nThe lateral acceleration ay can be measured using available sensors. When its kinematic equation is considered, it is observed that it consists of two components. The first one is the derivative of vy relative to vehicle fixed coordinate system and the second one is the acceleration component caused by the motion of the vehicle fixed coordinate system [12]. Under the assumption that the side\u2013slip angle \u03b2 is\nsmall, the lateral acceleration ay equation can be represented as\nay = v\u0307\u03b2 + v\u03b2\u0307 + v\u03c8\u0307 (6)\nIII. VEHICLE PARAMETER IDENTIFICATION\nFrom (1), the state variables \u03b2, \u03c8\u0307, \u03c6\u0307, \u03c6; the speed \u03c5 and the lateral acceleration ay are assumed to be measurable signals. The parameters m, g, Jxx, Jzz , L, T are assumed to be known with values indicated in Table I. The parameters l\u03c5 , C\u03c5 , Ch, k, c and h are considered as the unknown parameters to be estimated. The Recursive Least Squares (RLS) algorithm is used in [1] in order to estimate the lateral dynamics parameters in case when the speed is constant. However, the RLS technique does not perform well when the speed of the vehicle is time-varying. Therefore in this section we propose to use Kalman filtering techniques in vehicle parameter identification.\nBy taking the Laplace transform of the yaw\u2013rate equation\nand by letting\nc1 = Chlh \u2212 C\u03c5l\u03c5\nJzz\n, c2 = C\u03c5l2\u03c5 + Chl2h\nJzz\u03c5 , c3 = C\u03c5l\u03c5 Jzz\n\u2126l1(s) = \u03b2(s)\ns \u2212 \u03bb , \u2126l2(s) =\n\u03a8(s) s \u2212 \u03bb , \u2126l3(s) = \u2206(s) s \u2212 \u03bb\nwhere \u03bb is a positive parameter, we obtain\n\u00af\u0307\u03a8(s) = \u03a8\u0307(s) + T\n2Jzz\nU(s)\ns \u2212 \u03bb = c1\u2126l1(s) + (c2 \u2212 \u03bb)\u2126l2(s) + c3\u2126l3(s) (7)\nHence \u00af\u0307 \u03c8 can be represented as\n\u00af\u0307 \u03c8 = [\n\u03c9l1 \u03c9l2 \u03c9l3\n]\n\n\n\u03b81 \u03b82 \u03b83\n\n (8)\nAs time goes on, we have new measurements for \u00af\u0307 \u03c8 and filtered signals. If the speed is not constant during the estimation process, we can not use the least squares algorithm reliably. If the speed changes over time, the parameter \u03b82 which is the coefficient of the filtered signal \u03c9l2 is no more the same for all instants, i.e.,\n\u03b82 = \u2212 C\u03c5l2\u03c5 + Chl2h Jzz\u03c5 \u2212 \u03bb\nOn the other hand, by letting \u03b8 to be the state vector, the Kalman filter can be used. To this end, let\n\u03b82(k \u2212 1) = \u2212 C\u03c5l2\u03c5 + Chl2h Jzz\u03c5(k \u2212 1) \u2212 \u03bb (9)\nMultiply (9) by \u03c5(k \u2212 1)/\u03c5(k) and substitute the result into (9) to obtain\n\u03b82(k) = \u03c5(k \u2212 1)\n\u03c5(k) \u03b82(k \u2212 1) +\n\u03c5(k \u2212 1) \u2212 \u03c5(k)\n\u03c5(k) \u03bb (10)\n\u03b81 and \u03b83 are time invariant. If we consider \u03b8 vector as the state vector, \u00af\u0307 \u03c8 as the output and \u03bb as the constant input of"
        ]
    },
    {
        "image_filename": "designv11_100_0003922_2013-26-0143-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003922_2013-26-0143-Figure5-1.png",
        "caption": "Figure 5. Run Flat System and Run Flat Wheel Assembly.",
        "texts": [
            " Co-driver seat bottom \u2022 Overall length : 5746 mm \u2022 Width : 2160 mm \u2022 Height : 2850mm \u2022 Ground cle. : 311 mm \u2022 FAW : 3750 kg \u2022 RAW : 3900 kg \u2022 GVW : 7650 kg \u2022 Model : 6 BT- 5.9 TC \u2022 Type : 6-Cylinder, In-line \u2022 Max. Power : 120 hp @ 2500 rpm \u2022 Max. Torque : 410 Nm @1800 rpm \u2022 Allison Automatic transmission model 3060 Axles (ratio 5.857:1) \u2022 Front Axle : Salisbury Type \u2022 Rear Axle : Banjo Type \u2022 Semi elliptical at front and rear with shock absorber \u2022 Spring width(F) : 70 mm \u2022 Spring width(R) : 80 mm \u2022 Rim : 8.00 X 20 with Bias ply tire. \u2022 Rim : 8.25 X 22.5 with Radial Run Flat tire (Fig. 5) \u2022 Tire : 12X20 18PR Bias ply \u2022 Tire : 11R22.5, 16 PR Radial Run Flat \u2022 Front/Rear : 65 Details of High Mobility Vehicle (HMV) shown in Fig. 4, which is used for collection of test data are given below. Nomenclature : TATA LPTA 713/32 Confi guration : 4 X 4 Type of drive : Right Hand Drive The procedure for the objective assessment of ride quality has been carried out in accordance with ISO-2631-Part (1). The ride test consists of running vehicle over various Tracks at various speeds for objective measurement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002258_i2mtc.2013.6555379-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002258_i2mtc.2013.6555379-Figure6-1.png",
        "caption": "Figure 6. Spindle FEA. (a)Spindle CAD model (b)Boundary conditions (c)Spindle axial deformation because of 1N force aplied.",
        "texts": [
            " The components are considered as springs and therefore its axial deformation will depend of the applied force and the axial elastic constant (Eq. 1). 1 Where ki is the axial elastic constant of the component i and xi is the total displacement in the axial direction of the component i. For determining the axial elastic constant of a component, a fixed support boundary condition was placed at one end of the component, while at the other end a force of 1N is defined. The total displacement obtained with the FEA is used to calculate de rigidity of the part. Figure 6 shows an example of this procedure for the spindle. The total axial deformation of the spindle is 6.297\u00d710 -10 m. Therefore, the axial elastic constant of the spindle k23 is 1.588\u00d710 9 N/m. The same procedure was applied for determining the structural elastic constant of the other component. Transmission: The axial elastic constant of the gear and worm are determided as previously explained. Then, the equivalent gear stiffness related to the worm axis (kc-sf) is calculated as is shown in the Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.2-1.png",
        "caption": "Fig. 7.2 RR-type and RP-type 2-link robots",
        "texts": [
            " A further study on the theory of C-manifolds and their embeddings shows that for any two n-joint robot arms with a common structural configuration, their C-manifolds are topologically equivalent (homeomorphic). However, they may have different geometrical details due to the different sizes and dimensions in their mechanical structures. Therefore, it is clear that the topology of the C-manifold for a dynamic system is determined by its kinematics, while the geometry of the C-manifold is determined by its dynamics. The most simple but typical examples are the two-link robots: one has two revolute joints (RR-type) and the other one is of revolute-prismatic (RP) type, as shown in Figure 7.2. The C-manifold of the RR-type arm is homeomorphic to the surface of 2-torus T 2 S1 \u00d7 S1, while the second one has a C-manifold topologically equivalent to a 2D cylindrical surface S1 \u00d7 I1, as shown in Figure 7.3, where I1 = [0, 1] is the 1-dimensional unity interval. In addition, for each type of the two-link robot arms, a variation of the kinematic parameters, such as the link length and joint off-set, and a variation of the dynamic parameters, such as the mass, mass center, and the moment of inertia, will cause each individual C-manifold to have a different geometry, while the topology remains invariant"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003111_imece2012-86262-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003111_imece2012-86262-Figure3-1.png",
        "caption": "Fig. 3: Misalignment model for transition speed",
        "texts": [
            " Otherwise, they should be modified, and the solution procedure above is repeated until the force balance is reached. The transition speed is a speed at which the hydrodynamic bearing begins to raise and support the shaft and the rolling bearing release its support of the shaft. At the transition speed, the equilibrium positions locus of conventional hydrodynamic bearing intersects the clearance circle of the rolling bearing. Therefore, the transition speed can be determined according to the geometrical relationship of the locus and clearance circle. Figure 3 shows the model for calculation of the transition speed. In this figure, ec and \u03b8c are magnitude and direction of the misalignment, respectively. For a given misalignment, the transition speed can be solved as follows: First, the shaft speed U is set to an initial value; Then, the eccentricity and attitude angle of the hydrodynamic bearing are calculated according to Eqs. (6), and (9) to (11); The position Oe(xe,ze) of the shaft center at the initial shaft speed is expressed as: sin cos e e x e z e \u03b8 \u03b8 = \u2212  = \u2212 (16) The intersecting point of straight line OHOe and the clearance circle of the rolling bearing OJ(xJ,zJ) can be expressed as: ( ) ( ) ( ) ( )( ) 2 2 2 2 2 2 2 2 2 2 2 2 1 1 sin cos sin cos 1 J J c c c c c c c R a bx k k a b z k a e k b e k k e C \u03b8 \u03b8 \u03b8 \u03b8 \u2212 = + \u2212 = + = \u2212 + = + \u2212 + \u2212   (17) The distance between the Oe and OJ is ( ) ( )2 2 e J e J e JO O x x z z= \u2212 + \u2212 (18) The convergence criterion is expressed as: e JO O \u03b2\u2264 (19) where \u03b2 is an allowable relative error",
            "org/about-asme/terms-of-use 7 Fortunately, the maximum eccentricity ratio and the transition speed in the regions of \u03b8c near 140\u00b0 and 260\u00b0 just present slight fluctuations, and are almost same with design values under perfect alignment. One reason is that the deviations of the axis of rolling bearing from that of hydrodynamic bearing are compensated by contact deformations of rolling bearing; Another reason is that the location variations of the clearance circle of rolling bearing resulted in misalignment does not obviously change the position of intersecting point OJ (as shown in Fig. 3). To sum up, misalignment significantly affects the locus of equilibrium positions, the maximum eccentricity ratio, and the transition speed. However, regions of included angle \u03b8c insensitive to the maximum eccentricity ratio and the transition speed are presented. According to the insensitive regions, requirement of alignment errors can be made, and furthermore the machining and concentricity tolerance of the rolling bearing and hydrodynamic bearing can be determined. In this study, a misalignment model for rolling bearing and hydrodynamic bearing is developed to investigate the effect of the misalignment on performance of HRHB"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.16-1.png",
        "caption": "Fig. 2.16 4PRRPR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology P\\R\\R\\P\\kR",
        "texts": [
            "1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig. 2.1g0) Idem No. 13 34. 4RRRC (Fig. 2.19a) R\\R||R||C (Fig. 2.1h0) Idem No. 21 35. 4RCRR (Fig. 2.19b) R\\C||R||R (Fig. 2.1l0) Idem No. 21 36. 4RRCR (Fig. 2.20a) R\\R||C||R (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001573_978-1-4471-2343-9_19-Figure19.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001573_978-1-4471-2343-9_19-Figure19.5-1.png",
        "caption": "Fig. 19.5 Manipulator model: (a) top view, (b) side view in the case of q = 03",
        "texts": [],
        "surrounding_texts": [
            "The 3-dof planar manipulator is controlled by a PC running a real-time Linux OS. The Linux kernel is patched with Xenomai (http://www.xenomai.org/). The control period is 1ms. The rotating angle of each joint is obtained from an encoder attached to each DC servo motor via a counter board. Armature currents based on (19.11), (19.4) and (19.7) are interpolated to each DC servo motor by a D/A board through each DC servo driver. Here we consider the manipulator model as depicted in 19.5, where qi, \u03b1i, and di denote the i-th joint angle, the tilt angle, and the distance from the (i\u22121)-th joint to the i-th joint, respectively. Note that swc and q\u0307 are associated by the relation swc = cJ(q23)q\u0307 = \u23a1\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 d1C23 + d2C3 + d3 d2C3 + d3 d3 \u2212d1S\u03b13S23\u2212d2S\u03b13S3 \u2212d2S\u03b13S3 0 d1C\u03b13S23 + d2C\u03b13S3 d2C\u03b13S3 0 0 0 0 C\u03b13 C\u03b13 C\u03b13 S\u03b13 S\u03b13 S\u03b13 \u23a4\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 q\u0307, (19.12) where q23 := [q2, q3 ] , Si jk := sin(qi + q j + qk), Ci jk := cos(qi + q j + qk), S\u03b13 := sin\u03b13, and C\u03b13 := cos\u03b13, respectively. The physical and nominal parameters of the manipulator are as follows: (qmax 1 , qmax 2 , qmax 3 ) = (\u2212qmin 1 ,\u2212qmin 2 ,\u2212qmin 3 ) = (\u03c0/2rad, 5\u03c0/18rad, \u03c0/2rad) (= (90\u25e6, 50\u25e6, 90\u25e6 )), d1 = d2 = 0.2m, d3 = 0.047m, \u03b13 = \u03c0/6 (= 30\u25e6 ), M\u0304d = diag{0.65kg\u00b7m2, 0.25kg\u00b7m2, 0.06kg\u00b7m2}, K\u0304t = diag{21N\u00b7m/A, 17.64N\u00b7m/A, 4.96N\u00b7m/A}. The image resolution and the focal lengths of the CCD camera are 640\u00d7480 pixels, \u03bbx = 800.0 pixels, and \u03bby = 796.4 pixels, respectively. The frame rate of the camera is 120fps. Hence, the image data is updated every 8.3ms. The coordinate of the feature point is calculated as the barycentric coordinate of the marker area on the binarized image. The single camera system can not measure directly the depth on \u03a3c, i.e. czo1 . The matrix Jimg which consists Jvis depends on the depth czo1 , and therefore we need to estimate it. Similarly to our previous work in [7, 8], we collected some measured data of set (czo1 ,v1) in advance and fitted them to a second-order polynomial function with respect to v1. The depth czo1 is estimated by cz\u0302o1(v1) = av2 1 +bv1 + c, where a = 1.07348\u00d710\u22126, b =\u22126.24461\u00d710\u22124, c = 3.10071\u00d710\u22121. The procedure of the experiment is as follows: Step 1: We set qd =(\u2212\u03c0/4rad, 2\u03c0/9rad, \u03c0/36rad) (= (\u221245\u25e6, 40\u25e6, 5\u25e6)) and q\u0307d = 03 in (19.11) to drive the hand-eye robot to the initial configuration, so that the target object is inside the boundaries of the image plane and the 3-dof planar manipulator does not configure a kinematic singularity1. In consequence, the feature point is located around f = (140 pixels, 60 pixels). Step 2: We start visual servoing based on (19.11), (19.4), and (19.6) with fd = (0, 0). Step 3: When two seconds have passed, visual servoing is finished. The experiments were carried out with design parameters \u03c9 = 150rad/s, Kp = diag{144,144,144}, Kv = diag{48,48,48}, Kimg = diag{4,4}, \u03c1 = 0.01. The experimental results are shown in Figs. 19.6\u201319.8. For comparison, Figs. 19.7 and 19.8 show the results in the cases of the Zghal-type [18] and the Marchand-type [12] functions. These are typical performance functions for avoiding joint limits. Each graph includes the case of three kinds of kr and the case of no redundancy (V = 0). The two graphs in the second row of 19.6 show that the feature point f converged to fd , i.e. visual servoing as the primary task was achieved, at about 1s in every case. We here omit time responses of f in the case of the Zghal-type and Marchand-type functions, because these are similar to the case of the proposed tangential function. On the other hand, the results for joint limit avoidance are as follows: in the cases of the proposed and the Zghal-type functions, the joint limit avoidance succeeded without depending on the value of kr; in the case of the Marchand-type function, the second joint angle exceeded its limit depending on the value of kr. Focusing on the time response of q2 after f converged to fd , we can find a difference between the cases. In the cases of the proposed tangential function and the 1 Note that when (q2,q3) = (\u00b1k2\u03c0 rad,\u00b1k3\u03c0 rad), \u2200k2,k3 \u2208 Z, the manipulator configures kinematic singularities. Marchand-type function, q2 converges to the neighborhood of the braking angle and q\u0304max 2 , respectively. This means that q2 stays in N max 2 in both cases. Meanwhile, in the case of the Zghal-type function, q2 goes to zero, i.e. a kinematic singularity. This case needs avoidance of not only joint limits but also kinematic singularities. As described above, we confirmed that in IBVS the proposed tangential function can achieve both avoiding joint limits and exploiting joint range of motion effectively, i.e. admitting joint range of motion maximally. Therefore the effectiveness of the application was evaluated experimentally."
        ]
    },
    {
        "image_filename": "designv11_100_0001685_978-1-4302-4387-8_6-Figure3-4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001685_978-1-4302-4387-8_6-Figure3-4-1.png",
        "caption": "Figure 3-4. A custom wire-dispensing frame made from modular aluminum extrusion. This kind offixture",
        "texts": [
            " When testing wires, especially wires with attached connectors, it's a good idea to wiggle the wire about vigorously, as this will reveal some intermittent connections for you. If you're buying new wire, it will probably come on a reel. It's often handy to keep common sizes and colors on your bench, if you have room. A paper towel-holder makes an excellent wire dispenser rack, although you can also spend more time or money buying a prebuilt one or designing and fabricating your own. 63 CHAPTER 3 . COMPONENTS 64 A custom red -and -black wire dispenser is shown in Figure 3-4. You might use a fIxture like this if many pairs of identical-length wires are needed for a project. Not having to chase the wire spools allover the bench really saves a lot of time. This structure was built using a MicroRax miniature slotted aluminum extrusion (http://www.microrax.com), which is perhaps a bit of overkill, but was a lot of fun to plan and build. A coat hanger, artfully bent, would have also done the trick. makes it easier to cut both a red and a black wire to almost exactly the same length, which happens a lot when you use red and black wire for positive and negative power connections, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003882_icdma.2012.156-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003882_icdma.2012.156-Figure1-1.png",
        "caption": "Figure 1. GB12759-1991 the gear profile of basic rack in the normal plane",
        "texts": [
            " This paper has achieved the parametric modeling of three-dimensional entities of the double circular-arc gear by using CATIA parameters design, which lays a foundation for the double circular-arc gear strength analysis, simulation, virtual assembly design and development. II. THE ESTABLISHMENT OF THE DOUBLE The key of double circular-arc gear parameterized modeling is the establishment of the gear teeth curve profile. The double circular-arc gear teeth surface is enveloping formed during the process of the meshing movement of basic rack teeth surfaces while the he gear profile is formed point by point. Rack face teeth profile is the standard basic gear profile which is shown in figure 1. The double circulararc gear end-face teeth profile equations were deduced by the normal plane teeth shape of the basic rack. This type of GB12759-1991 double circular-arc gear, each teeth is constituted by eight circular arc. But the teeth profile formed by the standard rack is not circular arc but a curve similar to the circular arc. Derivation of the double circular-arc gear teeth surface equation[7]: 978-0-7695-4772-5/12 $26.00 \u00a9 2012 IEEE DOI 10.1109/ICDMA.2012.156 658 As shown in figure 2, when the plane of the rack section and the gear pitch cylinder roll with each other, rack teeth envelopes gear teeth surface"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001650_978-1-4419-8113-4_16-Figure16.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001650_978-1-4419-8113-4_16-Figure16.2-1.png",
        "caption": "Fig. 16.2 Bounce of a spinning ball dropped onto an inclined surface. If the surface is tilted to the right, the ball bounces almost vertically. If the surface is tilted to the left, the ball bounces a long way to the left",
        "texts": [
            " But if you try to rotate the ball around a vertical axis then the ball doesn\u2019t grip as well. The reason is that you can exert a torque by hand, on the edge of the ball, that is much larger than the torque exerted by the friction force near the axis. It is like loosening a nut with a long wrench, which is much easier than loosening the nut by hand. An important bounce event in baseball and softball, as well as in tennis and golf, is the bounce of a spinning ball off an inclined surface, as shown in Fig. 16.2. In tennis, a player can tilt the racquet head to vary both the rebound angle and spin of the ball. In fact, the modern game of tennis is dominated by the amount of spin that players impart to the ball. Players launch themselves off the court by belting the ball as hard as they can to spin the ball as fast as they can. Their opponent does the same, so the ball returns spinning furiously. If the player just taps or pushes the ball back at low speed, the ball will bounce off the strings at a strange angle",
            " In baseball and softball, the pitcher usually spins the ball rapidly so that it follows a strongly curved path through the air. The batter\u2019s main task is simply to connect with the ball, but if he is very good or just lucky, he can strike the ball above or below the axis to put even more spin on the ball. The ball impacts on the curved surface of the bat but the effect is the same as an impact on an inclined surface. The result is that the ball deflects skyward if the ball strikes above the long axis of the bat or it deflects down toward the ground if the ball strikes below the axis. In Fig. 16.2, the ball on the left is incident with backspin, while the ball on the right is incident with topspin. The direction of the spin is the same in both cases, but the ball on the left is incident from left to right relative to the surface, and the ball on the right is incident from right to left. If the ball was incident without any spin then it would bounce to the right off the left hand surface and to the left off the right hand surface. The effect of the counter-clockwise spin, on its own, is to deflect the ball to the left"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003942_amm.284-287.1799-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003942_amm.284-287.1799-Figure1-1.png",
        "caption": "Figure 1. Forces and moments on a quad-rotor.",
        "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. (ID: 130.88.90.140, The University of Manchester, Manchester, United Kingdom-22/05/15,15:54:47) The basic quad-rotor has symmetrical design and it consists of four complete rotors attached at equal distance from each other and a central hub. All the rotors are located within the same plane and oriented to generate thrust and torque as shown in Fig. 1. Each of the rotors on the quad-rotor helicopter produces both thrust and torque. Given that the front and rear motors both rotate counter-clockwise(make clockwise torque) and the other two rotate clockwise to balance the total torque of the system. The quad-rotor is controlled by separately adjusting the speed of the four rotors. Let Ti and \u03c4i be the thrust and torque for ith rotor respectively (i = 1,...,4). These values are normalized with the moment of inertia and mass, respectively. Denoting the distance of the rotor from the center of mass by l , we can introduce a set of four control input ui as function of normalized individual thrusts and torques as follows"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002529_mechatron.2011.5961097-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002529_mechatron.2011.5961097-Figure2-1.png",
        "caption": "Fig. 2. Schemes of multi-phalanxes grippers",
        "texts": [
            " Let us write this condition: \u0430 Hence: \u0430 It is necessary to consider here the possibility of contact of the second phalanx of the finger with the surface of the gripped element: Let us consider the influence of the multiple-link character of GD on the quality of gripping. The following versions of analysis according to variable parameters can be assigned here: the number of links; the length of links; and the combination of lengths of links of gripping fingers. Let\u2019s assume, that the phalanx of the finger stops pressing on the surface of the object, when the torque in the joint reaches the definite value. Then, for various numbers of phalanxes of GD we will obtain the following picture, see Fig.2. Analyzing the three-phalanxes gripper, one can conclude, that in general it will be ineffective, since at definite dimensions of the sphere (this range R\u2265a*sin\u03b1 is rather wide) one phalanx does not participate in fixation of the object, not contacting with its surface. However, even such gripping will be enough in order to fix safely and position accurately the manipulated object, if the pressure, provided by phalanxes, does not exceed the maximum allowable one. The scheme, shown in Fig. 3, is characterized by the presence of one long phalanx"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002355_j.jappmathmech.2012.03.010-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002355_j.jappmathmech.2012.03.010-Figure3-1.png",
        "caption": "Fig. 3.",
        "texts": [
            "8), in order to achieve uniform sliding of a punch over an elastic foundation a tangential force T has to be applied o it, exceeding the force P of Coulomb friction. When there is no friction ( = 0), this force is persists and takes the value T = = Ptg > 0. he paradoxical question arises: on what is the work of the difference between the forces T and P expended if there is no energy dissipation in n elastic body (irreversible deformations) ? A more paradoxical situation can be obtained if the case of a negative inclination of the punch base is considered (Fig. 3): g \u2019 (x) = \u2212 tg < 0. hen, instead of (1.8), the formula T = ( \u2212 tg )(1 + tg )\u22121P is obtained that, when < tg , gives T < 0, that is, in order to maintain uniform liding of the punch, a restraining force T, opposite to the sliding direction, has to be applied to it. In Fig. 3 we show how such a situation eads to the creation of a perpetual motion machine. . Formulae for the contact pressure Henceforth, the case when the elastic half-plane is taken as a foundation and there is no friction: q1 \u2261 0 is examined. With respect to the hape of the punch, it is assumed that g \u2019 (x) \u2208 H(\u2212 a, b), where H(\u2212a, b) is a class of functions satisfying the H\u00f6lder condition in the interval \u2212a, b) 8. When there is no friction, the contact pressure distribution is described by the integral equation9 (2",
            " What has been set out above enables us to draw the following conclusions. 1. When there is no friction between a punch and an elastic half-plane, a deformation force resisting the punch sliding arises that is aused by asymmetry in the punch shape and the contact pressure curve. The magnitude of the force is calculated using formula (3.4). I l f t A R 1 1 n the case of a smooth punch (Fig. 1) there is no force . However, in the case of a punch with end corners, this force can exist, and this eads to a violation of the law of conservation of energy (Fig. 3). 2. Analysis of the deformation of the half plane in the neighbourhoods of the end corners of a punch (Section 4) provides grounds in avour of the existence of corner tangential forces \u00b1 acting on the punch. The values of \u00b1, calculated using formula (4.2), compensate he deformation force formed at the limits of the contact area and thereby ensure that the law of conservation of energy is satisfied. cknowledgement This research was financed by the Russian Foundation for Basic Research (09-08-00901), 10-08-92001, 11-01-00650)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002414_amm.99-100.857-Figure4\uff0e3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002414_amm.99-100.857-Figure4\uff0e3-1.png",
        "caption": "Figure 4\uff0e3D Numerical Model of Journal Bearing and Operation Condition",
        "texts": [],
        "surrounding_texts": [
            "Numerical model of bearing and grid plotting The mathematics model of 3D journal bearing being studied in this paper is as figure 4 [14] shows: define the axis direction as X direction, the vertical direction as Y direction, the horizontal direction as Z direction. According to the estimation method of bearing\u2019s flow state, calculate the Re when the rotate speed reaches the biggest value, we can estimate the flow state is laminar flow when bearing\u2019s rotate speed reaches 500rad/s. Boundary condition The numerical model is same as the model in reference [8] , the shaft eccentric ratio of balance location is 0.42 when the load equals 30N. Different numbers of grid were tested to reach the grid independent. Table 2 shows the simulation results. Because of the pressure and stress is very less when the bearing moves in X direction, only the force of Y and Z direction are presented here. Comparing the numerical results of different grid number, we can get the conclusion that the number of grid has no influence to the result of numerical simulation when the number of grid reaches 998000. The comparison of different grid showed in table 2"
        ]
    },
    {
        "image_filename": "designv11_100_0001884_9781118609811.ch17-Figure17.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001884_9781118609811.ch17-Figure17.1-1.png",
        "caption": "Figure 17.1. Advantage of using silica-coated magnetic nanoparticles in enzyme immobilization.",
        "texts": [
            "26 Silica is often employed as a coating material for nanoparticles since it is chemically inert, promotes the dispersion of the nanoparticles, and has a high surface silanol concentration, which facilitates a wide variety of surface reactions and the binding of biomacromolecules. Furthermore, silica-coated magnetic nanoparticles are readily surface modifiable and can present a large number of functional groups, such as amino, aldehyde, and alkyl chain groups. For example, dendritic silica-coated magnetic nanoparticles were used to immobilize lipase and showed a higher enantioselectivity than that of free lipase (Figure 17.1).27 Porous Silica Nanostructures. Notwithstanding minimum diffusion limitations, enzyme loading per unit mass of nonporous nanoparticles is usually low. On the other hand, porous nanoparticles can afford high enzyme loadings due to their large surface areas. It should not come as a surprise that mesoporous silica particles have attracted a significant amount of attention as matrices for enzyme immobilization due to their high surface areas, controllable pore diameters, and uniform pore size distributions (Figure 17"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.31-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.31-1.png",
        "caption": "Figure 10.31 The flux distribution of an IPM line-start synchronous motor with circumferentialtype magnets: 1, magnet; 2, non-magnetic material; and 3, shaft",
        "texts": [
            " FEA is also cumbersome and time consuming in the early stages of PM motor design where numerous iterations are usually performed. Analytical calculation and analysis of all types of PM motors are essential in their early design stage. This section discusses the analytical method to calculate the air-gap flux of IPM machines using an equivalent magnetic circuit model taking into account the assembly gap and saturation in the steel. Factors that affect the flux leakage in an IPM motor will also be discussed. Figure 10.31 shows the configuration and no-load flux distribution of an eight-pole circumferential-type IPM line-start synchronous motor calculated using FEA. Integrated laminations are used to maintain integrity of the rotor. It contains three magnetic bridges in each pole: bridge I between the magnet and rotor slot; bridges II and III at the inter-polar space between the magnet and the shaft. Figure 10.32 shows the flux density along line I of Figure 10.31. It can be seen from Figure 10.32 that the flux densities differ in the two bridges. It will be shown later that the magnetic flux density in the magnetic bridges is related to the width and length of the magnetic bridge, rather than being constant. For the situation here, there is flux leakage in the rotor slot and the non-magnetic material between the magnet and the shaft. The flux leakage through the stator slot is negligible. Modern rare earth permanent magnets (REPMs) have a straight demagnetization curve as shown in Figure 10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.3-1.png",
        "caption": "Fig. 3.3 Double rod pendulum. (a) Pendulum angles and point masses moving in phase. (b) Pendulum angles and point masses moving out of phase. (c) Pendulum angles moving out of phase. Point mass m2 moving vertically",
        "texts": [
            " As an example, consider an idealised compound pendulum consisting of a rigid uniform rod, length L, with a frictionless pivot at one end (Fig. 3.2). The moment of inertia, I , is ML2/3, where M is the mass of the rod, the radius of gyration, , is L= p 3, and the distance of the centre of mass from the frictionless pivot, h, is L/2. Hence, from Eq. 3.4, the effective length, l , of the compound pendulum is 2L/3. A double rod pendulum consists of two simple rod pendulums (Fig. 2.1a) arranged in series, as shown in Fig. 3.3 (Lamb 1923; Kibble and Berkshire 1996; Baker and Blackburn 2005; Vanko\u0301 2007). The upper rigid massless rod, length l1, is suspended at its upper end from a horizontal frictionless pivot, and has a point mass, m1, at its lower end. The lower rigid massless rod, length l2, is suspended at its upper end from a horizontal frictionless pivot at the lower end of the upper rod, and has a point mass, m2, at its lower end. The two frictionless pivots are parallel so the motion of the point masses is confined to a vertical plane, and there are two degrees of freedom",
            " For small amplitudes a double rod pendulum has two modes of oscillation (Lamb 1923). The frictionless pivots are parallel so these are planar modes of oscillation. In both modes the lower rod oscillates about a virtual frictionless pivot. Which mode appears depends on how the pendulum is launched. Both modes of oscillation can be demonstrated by using the Meccano model. In the first mode of oscillation both the pendulum angles, 1 and 2, and the point masses m1 and m2 move in phase, that is the phase angle, , of both point masses at a given time is the same (Fig. 3.3a). Hence the phase, \". which is the difference between the phase angles is zero. The horizontal components of the displacements of m1 and m2, x1 and x2, have the same sign. In the second mode of oscillation the pendulum angles, 1 and 2 move out of phase, that is the phase angle is 180\u0131. For a given double rod pendulum the frequency of oscillation is the second mode of oscillation is always greater than in the first mode of oscillation. Geometric considerations show that, for the second mode of oscillation there are two possibilities for the point masses. In the first (Fig. 3.3b) they, and the horizontal displacements, x1 and x2, move out of phase so have opposite signs. The point mass m2 moves up and down at twice the pendulum frequency. The maximum values of the vertical displacement, y2, are at the ends of the arc followed by point mass m1. In the special case shown in Fig. 3.3c x2 is zero. The position of the virtual frictionless pivot of the lower rod is indeterminate, and behaviour can be chaotic. Lamb (1923) gives solutions for some special cases. For m1<< m2, in the first mode of oscillation 2 zero and the pendulum oscillates much like a simple rod pendulum of length (l1 C l2) The time of swing, T , is given approximately by (cf Eq. 2.13) T D 2 s l1 C l2 g (3.6) In the second mode of oscillation x1 x2 and y2 x2 that is the lower point mass, m2, is nearly stationary (Fig. 3.3c). For l1 l2 the upper point mass, m1, oscillates rather like a point mass m1, attached to a massless elastic string under a tension m2g, stretched between fixed points (Fig. 1.13). This is the point mass on a string analogy, and the time of swing, T , is given approximately by T D s m1 .l1 C l2/ m2g (3.7) which is much shorter than that given by Eq. 3.6. For m1 >> m2 and l1 l2 a double rod pendulum displays chaotic behaviour, as noted by Lamb (1923), who does not use that term. The double rod pendulum is one of the simplest mechanical systems that can display chaotic behaviour",
            " There are videos of chaotic behaviour, on YouTube, of both physical models of double pendulums and computer simulations (Anonymous 2010). Chaotic behaviour can be demonstrated by using the Meccano model. The appeal of a chaotic mechanical system is essentially visual, and the double rod pendulum is also an amusing toy. In general, when the Meccano model is launched, its motions are chaotic, with no apparent relationship between the motions of the bobs. However, as the motions decay, due to the effects of friction etc., the motions tend to degenerate into the first mode of oscillation (Fig. 3.3a). The two parts of a double rod pendulum can be regarded as coupled pendulums in which energy is transferred between pendulums. Coupled pendulums have a tendency to become synchronised, as first observed by Huygens in 1665 (Baker and Blackburn 2005), and recently discussed by Haine (2010). The discovery of the Blackburn pendulum, also called Blackburn\u2019s pendulum, is sometimes attributed to Hugh Blackburn, who first described it in 1844 (Benham 1909; Ashton 2003). However, it was discussed by James Dean in 1815 and analysed mathematically by Nathaniel Bowditch later in the same year (Whitaker 1991, 2005)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002982_amr.317-319.281-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002982_amr.317-319.281-Figure6-1.png",
        "caption": "Fig. 6. Stress distribution of flexspline Fig. 7. Stress distribution of flexspline gear ring",
        "texts": [],
        "surrounding_texts": [
            "This work is financially supported by the National Natural Science Foundation of China (Grant NO. 50975295)"
        ]
    },
    {
        "image_filename": "designv11_100_0002867_icssem.2011.6081325-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002867_icssem.2011.6081325-Figure3-1.png",
        "caption": "Figure 3. A distribution valve diagram.",
        "texts": [
            " BASED ON THE STRUCTURE IMPROVEMENT OF DISTRIBUTION VALVE REDUCE PRESSURE FLUCTUA nONS The backwash orifice shape and size can affect the switch-on time between the backwash pole and axis nozzle, which affects the backwash pressure fluctuations rate, and long switch-on time can leads to low pressure fluctuations rate while short switch-on time can leads to high pressure fluctuations rate. By using TT-45 (3 m 2 ) type ceramic filter as an example, the related parameters of the range of inlet pressure recoiling washing is 0.02-0.01 MPa. To prolong the switch-on time between the backwash pole and axis nozzle, we change the internal structure of the distribution valve from round to waist, which can reduce the rate of flow through the backwash orifice, and reduce pressure fluctuations. Figure 3 shows the internal diagram of the distribution valve. Figure 4 is the diagram for the distribution valve backwash orifice improvement. We design the inside round backwash orifice to waist, that is, extend original circular to waist on both sides, and expand the two waist holes for original vacuum orifice and suck orifice to large waist along two sides of circumference, which changes the backwash orifice area from 490.625 mm2 to 866.12mm2, while the other sizes don't change. With this improvement, the switch on time between the backwash pole and axis nozzle increases from original 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002567_icmtma.2011.343-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002567_icmtma.2011.343-Figure8-1.png",
        "caption": "Figure 8. The equivalent displacement map",
        "texts": [
            " For the ball, it has not only plastic deformation but also displacement relative to wire race. To ensure that the model is reasonable, for the two sections, by the symmetry of the model, we should constrain it\u2019s normal displacement having no any influences on tangential displacement. By the given load of platform, we applied surface load which is perpendicular to the ball section on 1/4 spherical. The value is 0.6N/ mm 2 .The model is shown in Figure 6. The equivalent stress map is shown in Figure 7 (a) and (b). Figure 7(b) is the local magnification map on contact area. Figure 8 shows the equivalent displacement map. We can see that the maximal contact stress of contact region is 100Mpa, the maximal displacement is 0.009mm and the contact stress of wire is significantly larger than that of ball. (1) In the course of bearing\u2019s practical work, wires are often damaged erlier than balls. The reason is that the yield limit and ultimate strength of 65Mn quenched spring steel are 800MPa and 1000MPa respectively, the minimum allowable safety factor is about 2. FEA results show that the maximal contact stress on contact area is 100Mpa, the actual safety factor of the maximal contact stress part is about 8, larger than the minimum allowable safety factor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.4-1.png",
        "caption": "Fig. 2.4 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRRR (a) and 4PRRRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3; xd\u00de, TF = 0, NF = 2, limb topology R||R\\R||R||R (a) and P\\R\\R||R||R (b)",
        "texts": [
            " PM type Limb topology Connecting conditions 1. 4RRRRR (Fig. 2.2a) R\\R||R||R\\k R (Fig. 2.1a) The first and the last revolute joints of the four limbs have parallel axes 2. 4RRRRR (Fig. 2.2b) R||R\\R||R\\R (Fig. 2.1b) The first, the second and the last revolute joints of the four limbs have parallel axes 3. 4RRRRR (Fig. 2.3a) R||R||R\\R||R (Fig. 2.1c) The two last revolute joints of the four limbs have parallel axes 4. 4RRRRR (Fig. 2.3b) R||R||R\\R||R (Fig. 2.1c) The three first revolute joints of the four limbs have parallel axes 5. 4RRRRR (Fig. 2.4a) R||R\\R||R||R (Fig. 2.1d) The two first revolute joints of the four limbs have parallel axes. 6. 4PRRRR (Fig. 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7. 4RRRPR (Fig. 2.5a) R||R\\R\\P\\kR (Fig. 2.1f) Idem No. 5 8. 4RRPRR (Fig. 2.5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10. 4PRRRR (Fig. 2.6b) P||R\\R||R\\R (Fig. 2.1i) Idem No. 9 11. 4RRPRR (Fig. 2.7a) R\\R\\P\\kR\\R (Fig. 2.1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002997_amr.452-453.1455-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002997_amr.452-453.1455-Figure1-1.png",
        "caption": "Fig. 1 (a) Front view of a low hysteresis brush seal, (b) section view of the geometry",
        "texts": [
            " The research has dealt with two primary issues: (i) the analysis and measurement of flow and gas leakage through the seal, and (ii) the analysis of brush seal force systems during service. A conventional brush seal includes a set of fine diameter fibers densely packed between the front and back plates. As a result of the bristle hysteresis, the bristles cannot follow the rotor excursion instantaneously[2], thus increase the gas leakage during operation. To minimize the hysteresis effect, a low hysteresis brush seal with a relief between the bristle pack and the back plate was developed[3], shown in Fig. 1(a) and (b). Bristles are aligned at lay angle \u03b8 , the radial clearance/interference is \u2206 , and it is assumed that there are N bristles along the z axis in each axial section. Aksit and Tichy [4] chose a discretization technique to compute bristle force and deformation by taking advantage of 3D quadratic beam elements. Their analysis includes bristle/rotor contact, inter-bristle contact and bristle/back platecontact. In a series of studies, Stango and H. Zhao [5.6.7] analyzed rotor/bristle contacts by (i) ignoring aerodynamic loads, (ii) taking into account radial aerodynamic loads but regarding them as constant ones, and (iii) integrating inter-bristle contact and aerodynamic loads as overall constant loads respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure15-1.png",
        "caption": "Figure 15. Normalised Stress at RSFB region (Contact analysis)",
        "texts": [],
        "surrounding_texts": [
            "Surface to surface contact is defined for the following contact pairs \u2022 RSFB bracket & External flitch \u2022 External Flitch & FSM \u2022 FSM & Internal flitch \u2022 Internal Flitch & RSFB Cross member"
        ]
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure23-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure23-1.png",
        "caption": "Figure 23 FDM candidate part design with text (see online version for colours)",
        "texts": [
            " The fixture must be capable of withstanding a 93\u00b0C sustained operating temperature and offer accuracy comparable to basic fixture tooling. The fixture geometry is shown in Figure 22. FDM was chosen due to the ability to build parts with high complexity quickly. A higher temperature capable material, PPSF, was chosen to accommodate the temperature requirements. In addition, due to the geometry flexibility, design engineers engraved text work instructions and quality inspection requirements directly into the computer-aided design model to aid test technicians in integrating the part in the test, as shown in Figure 23. To summarise, in order to justify RM from an economic standpoint, often design integration becomes a key element. Adding design integration as a critical gate when searching for AM part candidates exploits the fundamental nature of AM technology and design flexibility. Design integration coupled with economic analysis provides intelligent part candidate selection for AM technology. This evaluation of AM technology may be broken down into two distinct phases. Phase 1 includes the technical evaluation period where evaluating design integration opportunities is a critical element to the part candidate evaluation process"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure34-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure34-1.png",
        "caption": "Fig. 34 Chamfering Toe and Heel\u2014Large pitch cone angle spiral bevel gear",
        "texts": [
            " Therefore, the allowable amount of penetration Ds is a function of the current point on the tooth edge, the local spiral or helix angle, and the tip diameter and cone angle of the CT. One disadvantage of the Chamfer Tool is that it must attack the Toe or Heel tooth edges from outside the part, which means that for bevel gears with a small pitch cone angle, the Toe end of the teeth can be chamfered, and for gears with a larger pitch cone angle, the Heel end of the tooth can be chamfered. At the opposite end, it is likely a collision will result between the turn table and the tool spindle unless some adjustments are made to the operation parameters. For example (left, Fig. 34) the Chamfer Tool is set up for the Heel on a spiral bevel gear with a large pitch cone angle; it is clear that using a tool holder long enough allows avoiding any collision between the machine\u2019s tool spindle and the turn table. If the Toe end of the tooth is to be chamfered, the tool spindle clearly collides with the work and turn table (center, Fig. 34). The \u201cPivot Angle\u201d controls the angle between the Chamfer Tool\u2019s axis and the tooth Toe or Heel edge; in both the left and middle Fig. 34, the Pivot Angle is set to 90\u02da. If the Pivot Angle is now set to, say, 65\u02da, we rather get what appears in Fig. 34, right, and now, any risk of collision has disappeared and the chamfering operation can proceed. Figure 35 shows a typical chamfer at the Heel of a large pitch cone angle spiral bevel gear. The teeth are non generated. The chamfers on each flank show comparable sizes. However, at the root, the CT movement needs to continue deeper since a burr has remained. The main disadvantage of a non 90\u02da Pivot Angle is that the bottom part of the Chamfer Tool may become parallel to the bottom of the tooth or, even worse, its cutting edge may make a negative angle with the tooth gap root"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001271_s0167-8922(06)80016-7-Figure3.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001271_s0167-8922(06)80016-7-Figure3.2-1.png",
        "caption": "Fig. 3.2. Bearing test machine with electric current",
        "texts": [
            " Before making any measurement, the instrument was stabilized and the multiplier adjusted to get a convenient reading of the meter. The resistance (in megohm), R, of the test sample is given by: 28 Chapter 3 Behavior of Lubricants in Rolling-Element Bearings 29 R = Meter reading x Multiplier x Test Voltage 500 The resistivity of the test sample can be calculated by using the formula R x S c r - - ~ (3.1) cr is restivity, ohm cm, R is resistance, ohm, S is area of the electrodes cm 2, L is gap between electrodes cm. The Roller Bearing Test Machine, shown in Fig. 3.2, was used to evaluate the performance of rolling-element bearings rotating at speeds of up to 3000 rpm under different combinations of radial, horizontal and axial loads in static and dynamic modes of operation. Bearings of different sizes (of inner diameter ranging from 44 to 150 mm) can be tested, and investigations in various tribological areas can be programmed. The test machine is driven by a rectifier-controlled 17 kW DC motor, type AG 2503 AZ. The condition of the bearing under test is monitored by means of various thermocouples and vibration pick-ups mounted on the bearing outer race and the end-shield",
            " It is thus ensured that the current flows through the shaft, inner race of the bearing, rolling-elements, outer race of the bearing and the housing, to complete the electrical circuit. In the present studies, bearings type NU 326 lubricated with lithium base grease 'A' has been tested at 1100 rpm under 1000 kgf and 500 kgf of radial loads acting at two different directions (90 ~ to each other for a duration of 250 h by passing 50 A (AC)). During the test period voltage across the bearing varied from 1.12 to 2.3 V between points A and B (Fig. 3.2). Corrugated patterns on bearing surfaces and effect of operating parameters on the threshold voltages and impedance response were also studied. Similarly, NU 326 bearings were tested under identical conditions, without the passage of current under pure rolling friction. A Philips PW 1140 X-ray diffraction unit, complete with generator and goniometer, was used for analysis of soap residues of different greases. Instrument parameters are listed in Table 3.1. By X-ray diffraction, symmetry and regularity in arrangement of atoms was made visible when a monochromatic X-ray beam irradiates soap residues",
            " Infrared (IR) spectra of a sample of fresh grease 'A', and those of a sample of statically-bounded and electrically exposed grease 'A', taken after the completion of the eighth record, are shown in Figs. 3.8 and 3.9 respectively. Soap and oil were separated from these samples by dissolving them in petroleum ether. The IR spectra of soap and oil contents of these samples are shown in Figs. 3.10-3.13. Behavior of Lubricants in Rolling-Element Bearings 35 The vibration levels and temperature rise of the test bearing were periodically monitored during operation. The bearing impedance was monitored at the rated operating conditions; it was measured between points 'A' and 'B\" (Fig. 3.2). After 3.14-3.16 respectively. The bearing surfaces were found to have corroded after the passage of electric current, as shown in Fig. 3.17. The relative percentage of the lithium metal present in the sample was determined, and compared with those present in fresh and statically-bounded samples. Soap and oil contents of flesh grease and the grease samples from the test bearings and from motor bearings were separated by dissolving them in petroleum ether. The soap residues, thus obtained from different grease samples, were subjected B ehavior of L ubricants in R olling-E lem ent B earings "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003425_2506095.2506110-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003425_2506095.2506110-Figure1-1.png",
        "caption": "Figure 1. Frame arrangement of an AUV",
        "texts": [
            " MA and CA(v) are the added mass matrix and the added Coriolis and centripetal matrix respectively. D(v) is the damping matrix, respectively. )(g is the resultant vector of gravity and buoyancy.   T NMKZYX\u03c4  = Bf is the resultant input vector of thruster, control plane forces and moments.   T rqpwvu\u03bd  is the vector of linear and angular velocities in the vehicle coordinate frame.   T \u03c8\u03b8\u03c6zyx\u03b7  is the vector of absolute positions and Euler angles (roll, pitch and yaw). The relationship between linear and angular velocities in the vehicle frame to that in the absolute frame (refer Figure 1) is given by: \u03bd\u03b7J\u03b7 )( (3) where,  \u03b7J is the kinematic transformation matrix. For the controller design and closed-loop stability analysis, it is preferred that the system is investigated with respect to the earth fixed frame of reference in order to maintain every state to a single reference frame. For this, the coordinate transformation )()( \u03b7\u03b7,\u03b7,\u03bd   is performed using (3), which yields:                   \u03bd \u03b7 J(\u03b7(0 0I \u03b7 \u03b7  (4) The coordinate transformation \u03bc is a local diffeomorphism. This transformation is undefined for  =  90 and to overcome this singularity, a quaternion approach must be considered",
            " With u \u03b7g non-zero it may be possible to find continuous state feedback laws to stabilize the system as a whole. In addition to this, without loss of generality, we can assume that the damping terms of non-actuated states 22\u03b7D ~ are sufficiently large than their inertia terms 22\u03b7M ~ which means that the hydrodynamic restoring forces and torques are large enough to stabilize the non-actuated states (zero dynamics), which is a common property for AUVs [11]. This means that the robot can be exponentially stabilized by the actuated state controls alone. A flat-fish shaped AUV, as shown in Figure 1, is considered here for modeling and analysis. This AUV has a length of 4.5 m, width 1.46 m and depth 0.73 m. There are two propulsion thrusters for control of the longitudinal motion, four rudder planes and two stern planes for control of yaw and pitch respectively. The vehicle is under-actuated due to the absence of lateral (sway) and vertical (heave) thrusters. Because of underactuation, sway, heave and roll motion cannot be controlled. However, sway and heave motions can be achieved through the help of rudder and stern planes during trajectory tracking but not during zero speed control"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003619_itec.2013.6573488-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003619_itec.2013.6573488-Figure4-1.png",
        "caption": "Fig. 4. Symbolic representation of 5-phase induction motor",
        "texts": [
            " We recovers the formulation of the selection of the vector Vk+1, Vk+2, Vk-1, Vk-2, Vk+3, Vk+4 and Vk-3, well corresponding Vk-4 in a N=k zone, and this for a comparator two levels or three levels for the torque in case of changing the sense of rotation for the motor. III. MODELING OF THE SYSTEM FIVE-PHASE MOTOR \u2013 FIVE-LEG CONVERTER To make a study by numeric simulation of the DTC for the induction five-phase machine, we must establish the model of the machine in the Mark-reference () [4] [6] [7]. The induction machine proposed is with five-phases, four poles as shown on fig. 4. The saturation of iron is disregarded in the analysis. The basis equations of the motor are expressed in vectors of instantaneous shape. Tensions sV and the currents sI of the five-phase motor are represented, as follows [3] to [6]:  5 2 5 4 5 4 5 2 5 2  j e j d j c j bas evevevevvV    Equations of stator and rotor tension dt d iRV dt d iRV qs dsqssqs ds qsdssds       and dt d iRV dt d iRV qr drrqrrqr dr qrrdrrdr       )( )(  Equations of stator and rotor flux qrmqsmlsqs drmdsmlsds iLiLL iLiLL   )( )(   and qsmqrmlrqr dsmdrmlrdr iLiLL iLiLL   )( )(    Equation of the electromagnetic torque   dt d JCT ii p T mre dsqsqsdse     22 5 The analysis shows us that the main machine is the only one that produces the torque"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure38-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure38-1.png",
        "caption": "Fig. 38 Pinion Tip chamfering with Chamfer Tool",
        "texts": [],
        "surrounding_texts": [
            "direction. Such tools come in several base geometries, and tool manufacturers such as Sandvik, P. Horn, Iscar, Ingersoll Rand make disk bodies to different diameters on which blade inserts are screwed. The CoSIMT could also be a grinding disk. These tools can be used to chamfer pretty much any gear type, given the blade angles are chosen correctly. For example, the tooth tips of a spur gear can be chamfered using a CoSIMT with 15\u02da blade angles (left, Fig. 40) or 0\u02da (right, Fig. 40) and both will give the same results. The same comment applies to helical gears (left, Fig. 41), where the cutting edges are angled relative to the CoSIMT radius, and straight bevel gears (right, Fig. 41) where the cutting edges are parallel to the CoSIMT radius."
        ]
    },
    {
        "image_filename": "designv11_100_0003156_1.4005598-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003156_1.4005598-Figure3-1.png",
        "caption": "Fig. 3 The contour of the groove bottom at a set position",
        "texts": [
            " The milling process of the spiral groove is described using four right-handed Cartesian coordinates. Coordinate systems S1(X1, Y1, Z1) and S3(X3, Y3, Z3) describe starting positions of the cutter and the workpiece. Coordinate system S2(X2, Y2, Z2) denotes the rotation of the cutter around Z1-axis (located on the axis of the star-wheel) and the rotation of the workpiece around Z3-axis is given by coordinate system S4(X4, Y4, Z4). To get the groove bottom contour, a radial ray OA is assumed to be fixed on the workpiece as shown in Fig. 3. It rotates with the rotation of the workpiece. If the radial ray passes through the groove bottom, it will certainly intersect with the end face of the cutter, and the position of the intersection point could be calculated. When the workpiece and the cutter rotate, all of the intersection points constitute a surface line on the cutter end as shown by line 12 in Fig. 3. In the simulation, radial distance, q, from every intersection point to the axis of the workpiece is recorded. Among them, the minimum distance, qmin, could be found. This minimum distance denotes the point on the contour of the groove bottom, as noted by point C in Fig. 3. Furthermore, position of point C on the end face denotes position of a contact point to the groove bottom. By this way, the groove bottom contour at the set position of radial ray OA is obtained by finding the minimum distance, qmin, from axis of the workpiece to the points of intersection. Thus, points on the contour of the groove bottom at different positions could be obtained gradually by changing the position of radial ray OA. To calculate the intersection points between the radial ray and the end face of the cutter, mathematical models of the cutter and the workpiece are established",
            "org on 12/19/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use Here, /sr and /sw are rotation angles of the cutter and the workpiece. According to the motion relation between them, we can obtain the following equation: /sr \u00bc P/sw (4) where P is the tooth number ratio of the star-wheel to the screw rotor. Point B could be thus described in coordinate system S4 as RB4 \u00bc M42RB1 (5) 3.2 Mathematical Model of the Radial Ray. Radial ray OA is set fixed with the workpiece. Its position is described by f and u as shown in Fig. 3. Any point C on radial ray OA could be described as RC4 \u00bc q sin u q cos u f 1 2 664 3 775 (6) Here, f is Z4-axis coordinate of radial ray OA and u is the angle between radial ray OA and the plane Z4O4Y4. 3.3 Intersection Points of the Radial Ray to the End Face of the Cutter. To calculate the intersection point of radial ray OA to the end face of the cutter, we can assume that the arbitrary point B is the intersection point and it coincides with point C on the radial ray OA. According to Eqs. (2), (5), and (6), we get q sin u \u00bc a sin /sr l sin /sr cos\u00f0/sw \u00fe b\u00de r sin /sr cos h sin\u00f0/sw \u00fe b\u00de \u00fe \u00f0n\u00fe r sin h\u00de cos /sr m sin /sr sin /sw q cos u \u00bc a cos /sr l cos /sr cos\u00f0/sw \u00fe b\u00de r cos /sr cos h sin\u00f0/sw \u00fe b\u00de \u00f0n\u00fe r sin h\u00de sin /sr m cos /sr sin /sw f \u00bc \u00f0m l sin b\u00fe r cos h cos b\u00de cos /sw \u00f0l cos b\u00fe r cos h sin b\u00de sin /sw 8>< >: (7) Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003344_j.proeng.2012.04.115-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003344_j.proeng.2012.04.115-Figure1-1.png",
        "caption": "Fig. 1. Normal force and rolling resistance Fig. 2. Contact surface and tangential force of friction",
        "texts": [
            " In this approach the normal force of reaction between two solids in contact is given by the following expression: 1n v cnF k V av (1) where V is the volume of interpenetration, vcn is the velocity of the volumetric centroid of measured perpendicular to the contact patch, \u2018a\u2019 is the coefficient of damping and kv is the volumetric stiffness coefficient. For our analysis, we have assumed the cricket ball to be a perfect sphere. However, for nonspherical shapes, numerical or analytical integration can be used to compute the volume of penetration in an efficient fashion. Since we have assumed that both the ball and the pitch are deformable, the volumetric stiffness coefficient is defined in terms of the individual stiffness coefficients as v vBALL vGROUND vBALL vGROUNDk k k k k (2) Figure 1 shows the diagram for the evaluation of the normal force. The shaded area is the volume of interpenetration and the dotted line shows an exaggeration of the shape of the ball during impact. The direction of the contact normal is dependent on the properties of the penetrated volume but for the interaction between a cricket ball and a level pitch, it is assumed to be aligned with the upward direction. Since the reaction forces are dependent on the normal velocity, it can be clearly seen that these forces are not distributed evenly around the centroid of the contact volume. In figure 1, due to the angular velocity the normal velocity of the volume element \u2018B\u2019 is greater than that of element \u2018A\u2019. This makes the normal force acting at \u2018B\u2019 greater than that acting at \u2018A\u2019. This asymmetric distribution of normal forces gives rise to a net moment around the centroid \u2018c\u2019 opposing the angular velocity Gonthier 0 has demonstrated that the rolling resisting torques act around directions perpendicular to the contact normal and for the present example can be expressed as s v x v x x v x x s v z v z z v z z k a J k a J k a J k a J (3) In the above equation, and are the x and z components of the surface inertia tensor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001253_robot.2005.1570279-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001253_robot.2005.1570279-Figure6-1.png",
        "caption": "Fig. 6. Joints of the example robot arm and hand.",
        "texts": [
            " The gripper has 4 DOF: 3 of them translational (opening and closing of the fingers) and one rotational (abduction of the two articulated fingers, which are coupled). As it can be seen in the figure, multiple different grasps can be performed with such gripper, given three contact points on the object surface. First, three different finger assignments can be chosen (rows of the figure) and then, for each assignment it is possible to reach the contact points with different abduction angles (columns of the figure). The gripper is supposed to be attached to a SCARA robot arm, which in a 2D representation adds 3 DOF. Fig. 6 shows the planar structure of arm and gripper. The joints are numbered from 1 to 7, following the hierarchy order chosen for the application of the algorithm. The more relevant joint has been considered the one that controls finger abduction (only one joint as abduction is coupled in both articulated fingers). Even though there are other joints closer to the contact points, this is the one that better represents the kind of grasp being performed. Next, the translational gripper joints: no difference in relevance is found in them, so they have been randomly ordered"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002718_amm.197.46-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002718_amm.197.46-Figure2-1.png",
        "caption": "Figure 2. The finite element model of the positioning platform with connecting plate",
        "texts": [
            " Release X DOF and couple the DOF in other directions on the surface which is between X axial slider and the guiding rails on baseboard. Release Y axial DOF and couple the DOF in other directions on the surface which and couple the DOF in other directions on the surface which is between front shafting bearing and front lapping plate, back shafting bearing and back lapping plate. As a result, the finite element model of the positioning platform with connecting plate is established, and it is shown in Fig. 2. The natural frequencies of the XY positioning platform are gained through modal analysis, and they are summarized in Table 1. The first 6 th modal shapes of the positioning platform are shown in Figure 3. As shown in Fig.3, at the frequency of 165Hz, the X axial and Y axial tables driven by X axial motor and Y axial connecting plate driven by Y axial motor generate torsion vibration around the X axis in the flexure hinges of the elastic decoupling mechanism.. At the frequency of 184Hz, X axial and Y axial positioning platform and the Y axial connecting plate generate torsion vibration around the X axis in the flexure hinges of the elastic decoupling mechanism, while a torsion vibration around the Z axis at the same time"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure42-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure42-1.png",
        "caption": "Fig. 42 Spiral bevel gear Tip chamfering with a CoSIMT",
        "texts": [
            " 40) and both will give the same results. The same comment applies to helical gears (left, Fig. 41), where the cutting edges are angled relative to the CoSIMT radius, and straight bevel gears (right, Fig. 41) where the cutting edges are parallel to the CoSIMT radius. However, for spiral bevel gears, the CoSIMTmust have a convex side that is used to chamfer the concave tooth tip, and normally a flat or concave side that is used to chamfer the convex tooth tip, although a convex blade will also work as Fig. 42 shows. Four unit vectors (Fig. 43) are defined to orient the CoSIMT at any point along the tooth Tip edge of a spur or straight bevel gear: the cutting blades work away from their tips; N : the local normal vector,\u2212\u2192 Vo: unit vector along the tooth tip edge,\u2212\u2212\u2192 Tool: vector parallel to the tool axis. \u2212\u2192 Vo is obtained from the local partial derivative along the tooth tip. Vector N is pivoted about \u2212\u2192 Vo by half the difference between the local pressure angle and 90\u02da to yield the Gear Tooth Edge Deburring and Chamfering \u2026 179 tool axis vector \u2212\u2212\u2192 Tool"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001531_s00707-013-0995-y-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001531_s00707-013-0995-y-Figure3-1.png",
        "caption": "Fig. 3 Rigid beam in a moving reference frame",
        "texts": [],
        "surrounding_texts": [
            "conditions by successively degenerating the derivation grade, once more changing sign for odd derivatives. (This result reflects the procedure of integrating by parts, see also Eq. (33), with its consecutive sign changing). For the present example, the procedure is demonstrated in Table 1: the first row indicates the initial terms, i.e., D and dQc = [dQc/dx]dx , and the first column shows the resulting operators that are applied to [dQc/dx]. The result is displayed in the second column.\n2.1.1 Summary procedure\nThe foregoing is not restricted to only one spatial variable nor to one only deflection function (e.g., a plate w(x, y, t): one function, two spatial variables, or a beam qT = [\u03d1(x, t), v(x, t), w(x, t)]: one torsional and two bending functions, one spatial variable, see [7]). The latter case is summarized in Table 2.\n2.2 Elastic deflection suppressed\nWhen disregarding elastic deflections, we are left with\ny\u0307 = ( vox voy \u03c9oz )T (12)",
            "which characterizes the plane motion of a rigid body. The corresponding virtual work then reads\u222b B \u03b4yT[dM11y\u0308 + dG11y\u0307 \u2212 dQ1] = 0 with Eq. (6) et seq.\n\u222b B \u03b4yT \u23a7\u23a8 \u23a9 \u23a1 \u23a3 dm 0 \u2212dm v 0 dm dm x \u2212dm v dm x dJ o \u23a4 \u23a6 y\u0308+ \u23a1 \u23a3 0 \u2212dm\u03c9oz \u2212dm x\u03c9oz dm\u03c9oz 0 \u2212dm v\u03c9oz dm x\u03c9oz dm v\u03c9oz 0 \u23a4 \u23a6 y\u0307\u2212dQ1 \u23ab\u23ac \u23ad=0. (13)\nSince v = 0, spatial integration yields\n\u03b4yT \u23a7\u23a8 \u23a9 \u23a1 \u23a3m 0 0\n0 m m c 0 m c J o\n\u23a4 \u23a6 y\u0308 + \u23a1 \u23a3 0 \u2212m\u03c9oz \u2212mc\u03c9oz\nm\u03c9oz 0 0 mc \u03c9oz 0 0\n\u23a4 \u23a6 y\u0307 \u2212 Q \u23ab\u23ac \u23ad = 0\n:= \u03b4yT[My\u0308 + Gy\u0307 \u2212 Q] (14)\nwhere c: mass center distance.\n3 Multi-beam systems\n3.1 Rigid beam systems\nEquation (14) is now used to calculate a system of rigid bodies. This is easily done by summing up their individual virtual work according to Eq. (14):\nN\u2211 k=1 \u03b4yT k [ M y\u0308 + G y\u0307 \u2212Q ] k = 0 = \u03b4qT N\u2211 k=1\n( \u2202 y\u0307k\n\u2202q\u0307\n)T [ My\u0308 + Gy\u0307 \u2212 Q ] k . (15)\nThe mass center velocities of frame origin (i) are those of frame (p) at the coupling point L (p: predecessor of i), which are to be transformed into the actual frame and superimposed with the relative velocities of frame (i).",
            "and henceforth called the \u201ckinematic chain.\u201d Starting with i = 1, which does not have a predecessor, one obtains\n\u239b \u239c\u239c\u239d y\u03071 y\u03072 ...\ny\u0307N\n\u239e \u239f\u239f\u23a0 = \u23a1 \u23a2\u23a2\u23a3 E T21 E ... ... . . .\nTN1 TN2 \u00b7 \u00b7 \u00b7 E\n\u23a4 \u23a5\u23a5\u23a6 \u239b \u239c\u239c\u239d y\u03071r y\u03072r ...\ny\u0307Nr\n\u239e \u239f\u239f\u23a0 (18)\nwhere Ti j = Ti p(i) \u00d7 \u00b7 \u00b7 \u00b7 Ts( j) j (p: predecessor, s: successor). Next, the relative velocities are computed with y\u0307i,rel = Fi q\u0307i (Fi : local functional matrices) yielding\n\u239b \u239c\u239c\u239d y\u03071 y\u03072 ...\ny\u0307N\n\u239e \u239f\u239f\u23a0 = \u23a1 \u23a2\u23a2\u23a3 F1 T21F1 F2 ... ... . . .\nTN1F1 TN2F2 \u00b7 \u00b7 \u00b7 FN\n\u23a4 \u23a5\u23a5\u23a6 \u239b \u239c\u239c\u239d q\u03071 q\u03072 ...\nq\u0307N\n\u239e \u239f\u239f\u23a0 . (19)\nEquation (15) reads, in vector notation,\n[( \u2202 y\u03071\n\u2202q\u0307\n)T ( \u2202 y\u03072\n\u2202q\u0307\n)T \u00b7 \u00b7 \u00b7 ( \u2202 y\u0307N\n\u2202q\u0307 )T ]\u23a1\u23a2\u23a2\u23a2\u23a3 M1y\u03081 + G1y\u03071 \u2212 Q1 M2y\u03082 + G2y\u03072 \u2212 Q2 ...\nMN y\u0308N + GN y\u0307N \u2212 QN\n\u23a4 \u23a5\u23a5\u23a5\u23a6 = 0 (20)\nwhere q\u0307T := (q\u0307T 1 , q\u0307T 2 . . .) is defined with Eq. (19).\n3.1.1 Summary procedure\nInserting Eq. (19) into Eq. (20) yields a \u201cGaussian form\u201d according to Table 3. All non-marked elements are zero. Disposing thus of a representation with an upper (block-) triangular matrix enables to solve the motion equations recursively without inversion of the (global) mass matrix."
        ]
    },
    {
        "image_filename": "designv11_100_0002783_1.3543587-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002783_1.3543587-Figure3-1.png",
        "caption": "Fig. 3. Side view of single-spring hybrid oscillator.",
        "texts": [
            "5 Experimental values for the effective torsion constant and effective spring constant can be found by graphing tnet = \u2013keff q using T1R1 \u2013 T2R2 = \u2013 keff q with data from the force sensors and rotary motion sensor (see Fig. 2). Once keff is found, the relationship [see Eqs. (2) and (5)]eff eff 2 D k R \u03ba = can be used to find the effective spring constant. Torsion oscillator with one spring and a suspended mass A similar oscillating system can be created using only one spring if the other spring is replaced by a suspended mass exerting a torque on the disk (see Fig. 3). The effective spring and torsion constants of this hybrid oscillator can again be expressed using known parameters including the stiffness of the spring, the amount of hanging mass, the radii on which the spring and weight of the mass exert their forces, and the moment of inertia of the disk. As before, the spring is attached horizontally to the disk with a string that exerts force Ts perpendicularly to the radius on which the torque is exerted. The mass is suspended be- Equation (5) describes the system as a translational oscillator, where mD,eff and RD are, respectively, the effective mass and radius of the disk"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.9-1.png",
        "caption": "Figure 4.9 Design of bearing boss",
        "texts": [
            " Reducers of serial production are desirable to be designed so that the ends of the input and output shafts can be on each side. To achieve this, it is necessary for the bosses to be of the same dimensions. The width of the bearing location depends on the width of the actual gear and the sizes of elements for its fixing up. It should be taken into account that the bearings are built in bosses at a distance of 3\u20135mm from the internal edge of the housing wall if an adjusting ring does not exist (Figure 4.9). If this ring exists, the position of the bearing is determined by the actual ring width, that is, the width of the seal going inside the housing for 2\u20133mm. In order to increase stiffness and strength and for better heat withdrawal, especially in cast housings, ribs are applied. By the distribution of ribs, it is possible to improve casting conditions, facilitate contraction and reduce the occurrence of residual stresses. For a more uniform cross-section and better casting conditions, the common rib design is carried out pursuant to Figure 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure11-1.png",
        "caption": "Figure 11. Components in RSFB region",
        "texts": [],
        "surrounding_texts": [
            "Surface to surface contact is defined for the following contact pairs \u2022 RSFB bracket & External flitch \u2022 External Flitch & FSM \u2022 FSM & Internal flitch \u2022 Internal Flitch & RSFB Cross member"
        ]
    },
    {
        "image_filename": "designv11_100_0003749_robio.2012.6491229-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003749_robio.2012.6491229-Figure3-1.png",
        "caption": "Fig. 3. Throwing task. The robot transfers the object on the work-plane from initial position to throwing position and throws it toward goal position.",
        "texts": [
            " Dynamic Analysis If we assume that the three locations of initiation, throwing and target are specified (Initial location and target are inputs and the location of throwing is specified by the planner), the first thing that must be calculated is the linear and angular speed of the object at the point of throwing. Suppose that the goal location is denoted by , start location by , throw location by , and the coefficient of the friction between the object and the plane by . The plane can be inclined in general. We define global frame out of the plane. In Fig. 3 one sample of throwing task can be seen. In [10] a rather complete analysis of the forces exerted to the object by the environment at time of grasping the object with robot fingers is given. In [10], the distribution of the pressure of the object's surface on the plane is considered to be non-uniform. Here we consider the forces exerted to the object after release of the object from robot fingers, thus we assume a uniformly distributed friction force exerted to the object. We use energy analysis method to compute the velocity of the object at the point of throw: (2) In the foregoing equation, and are kinetic and potential energies of the object at the point of throw, and are kinetic and potential energies at the goal point and is the lost energy which equals the work of frictional force"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.5-1.png",
        "caption": "Fig. 3.5 Blackburn pendulum",
        "texts": [
            " The discovery of the Blackburn pendulum, also called Blackburn\u2019s pendulum, is sometimes attributed to Hugh Blackburn, who first described it in 1844 (Benham 1909; Ashton 2003). However, it was discussed by James Dean in 1815 and analysed mathematically by Nathaniel Bowditch later in the same year (Whitaker 1991, 2005). It appears to have been re-discovered independently by Blackburn. A Blackburn pendulum is a string pendulum, (Lamb 1923; Whitaker 1991, 2005; Baker and Blackburn 2005) with three inextensible, massless strings arranged in a Y shape (Fig. 3.5). The upper ends of the Y are clamped, and a point mass. m, is attached to the lower end. The distance from a line through the clamps to the centre of the Y is l1, and the distance from the centre of the Y to the point mass, m, is l2. It is assumed that the forces on the point mass are such that the forces in all three strings are tensile. Otherwise, a Blackburn pendulum does not maintain the configuration described above. A different type of pendulum, invented by J A Blackburn (Baker and Gollub 1996), and used for studying chaotic behaviour, is also known as the Blackburn pendulum",
            " Two orthogonal (at right angles) modes of oscillation are possible, so it has two degrees of freedom. In the first mode of oscillation the Y oscillates as a whole about the clamps and, for small amplitudes, the pendulum is equivalent to a simple rod pendulum (Fig. 2.1a) length l1 C l2 so the approximate time of swing, T , is given by Eq. 3.6 where g is the acceleration due to gravity. In the second mode the lower string oscillates about the centre of the Y in the plane of the Y (the plane of the paper in Fig. 3.5) and the approximate time of swing is given by T D 2 s l2 g (3.8) More generally, combinations of the two modes of oscillation are possible and the point mass, m2, moves on the surface of a torus generated by the rotation of a circle with centre at the centre of the Y and radius l2 about the line joining the clamps. In a bifilar pendulum (Gauld 2005) a point mass, m, is suspended from a pair of inextensible massless strings of equal length arranged so that the path of the point mass, m, is constrained to a vertical plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002350_09507116.2012.715919-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002350_09507116.2012.715919-Figure1-1.png",
        "caption": "Figure 1. Thermodynamic treatment of the heated surface of the component: 1, the torch; 2, the cooling system (the nozzle for supplying the lubricating cooling liquid); 3, the cutting tool; 4, the roller and 4, the components.",
        "texts": [
            " Surface-plastic deformation (SPD) was carried out by water cooling with a roller with the regulated pressure on the treated surface. The extent of heating of the surface prior to SPD was varied in a wide range by varying the heating time and the distance from the hightemperature core of the flame of the torch to the heated surface. Equipment for the investigation of the process of cutting with heating and SPD consists of a lathe with a suspended device. The experimental equipment is shown schematically in Figure 1. The blank was secured in the lathe in a three-jaw chuck with compression by the centre of the rear chuck. The standard acetylene\u2013oxygen torch is fitted with a reduction valve in which the gas flow rate is automatically regulated. The torch is assembled on a longitudinal support of the lathe in front of the cutting tool on the side opposite to the component. The roller is placed on the same longitudinal side of the lathe behind the cutting tool. The roller is set to the required pressure conditions using a screw pair and a built-in dynamometer"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003187_amm.483.382-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003187_amm.483.382-Figure7-1.png",
        "caption": "Figure 7 Schematic diagram of the bearing location Figure 8 Three-dimensional model diagram",
        "texts": [
            " The overall rotation mechanism with multiple brushes can circumferentially sweep the dirt in the wire rope gap. 3 is the big belt wheel, 4 is support, 5 is drawer, 6 is motor. Design and calculation of the transmission system of rotating brush (1) Motor selection Center of gravity is located in the center line. G=700N. Equation (3.1) can calculate the position of the center of gravity. i ii c F xF x \u03a3 \u03a3 = (3.1) The relationship between the center of gravity and the anchor is shown in Figure 6. F1=377N, F2=323N. Bearing arrangement is shown in Figure 7. Motor power is 0.77kW known by the formula 3.2. Taking into account friction and preload\uff0cSelecting the power of motor is 2.2kW, the model is Y112M-6-B3. (2) According to the actual condition, the speed of large pulley is 300r/min, the speed of small pulley is 1000r/min, rated power is 2.2kW. Accordance with the relevant calculation shows, the small pulley diameter is 75mm, the big belt pulley diameter is 250mm, the reference length of belt is 1600mm, the number of belt is 4. Finally, we can get the three-dimensional model of The Decontaminating Device"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.20-1.png",
        "caption": "Figure 4.20 Scheme of spray lubrication",
        "texts": [
            " Then the protective ring is placed in the inside of the housing and the cover on the outer side to avoid loss of lubricant. One-third to one-half of the free space of the bearing is then filled with grease if the rotational speed is above 1500min 1. The grease must be supplemented every three months and replaced with new grease once a year, after the bearing had been previously washed, cleaned and dried. Spray lubrication is applied at higher peripheral speeds. The scheme of the device is shown in Figure 4.20. It consists of a tank of oil (1) in which the oil flows from the gear housing (2) and the housings of other devices in the facility. From the tank through the suction box (3) with a strainer, the main pump (4) sucks oil and, when cooled in a cooler (5), pumps it up to each gear pair, to each bearing and to other places in the drive which require lubrication. A separate, smaller pump (6) also sucks oil from the tank and compresses it through a filter (7) back into the tank. In this way, clean oil is provided for the system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001202_tec.2021.3058804-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001202_tec.2021.3058804-Figure3-1.png",
        "caption": "Fig. 3. Six segmentations of the rotor bar.",
        "texts": [
            " The proposed model is evaluated using the four operating points listed in Table I, where Is denotes peak stator current, and fs, fe denote slip and synchronous frequency, respectively. The low and high values of stator current are selected to validate the model at different saturation levels. For each stator current, we have low and high values of fe at fixed arbitrary fs, to demonstrate the impact of increased rotor speed on cage loss. The loss is calculated per unit axial length. The sensitivity of loss with respect to the segmentation is studied using the parameters of Table II (see Section II-A of Part I). These six segmentations are illustrated in Fig. 3. Note that segmentation (f) has the same radial partitioning as (d) but lacks segments along the tangential direction. Unless otherwise mentioned, all results shown are for operating point (iv) with segmentation type (b). For the convergence of the fixed-point iteration in Subprob- lem A, we use the following tolerance values for (5), (6): \u01ebmag(\u03bd,m) = { 0.01 for \u03bd = 1, \u2200 m 0.05 for \u03bd = 3, 5, 7, \u2200 m, (30) \u01ebph(\u03bd,m) = { 0.05\u00b0 for \u03bd = 1, \u2200 m 2\u00b0 for \u03bd = 3, 5, 7, \u2200 m. (31) 2The stacking factor does not affect the proposed model accuracy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure16-1.png",
        "caption": "Figure 16. Input Force to the Articulation Sternoclavicularis and the Scapulothoracic Joint",
        "texts": [
            " The force of shoulder muscle, and the force of clavicle and the rib are shown in Fig. 13. B. Input to Scapulothoracic Joint (Analysis 2) External input c F is added at the scapulothoracic joint (Fig. 14 and Table VI). The force of shoulder muscle, the force of clavicle and the rib are shown in Fig. 15. C. Input to Articulatio Acromioclavicularis and Scapulothoracic Joint (Analysis 3) External input is added evenly at the articulation sternoclavicularis ( b F ) and the scapulothoracic joint ( c F ) (Fig. 16 and Table VII). The force of clavicle and the rib are shown in Fig. 17. D. Discussion In order to minimize the burden on the shoulder muscles, we compared the amount of force exercised by the muscles in the analysis 1 to 3 by integrating the absolute values of muscles forces 1 F , 2 F , 3 F , 4 F , when 2 \u03b8 is varied from 80 degrees to 110 degrees, as this is supposed to be the range most used. As shown, the value for the analysis 3 is the smallest. Thus, we can conclude that giving input force to two points is the most desirable method for training"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.17-1.png",
        "caption": "Fig. 3.17",
        "texts": [
            " 150 3 Dynamics of Rigid Bodies a l r a dA a b c a c dm r c z x C A r\u0304 y rc a yc xc x\u0304 y y x x x\u0304 y\u0304 r y\u0304 A dm rc C r\u0304 Fig. 3.16 The first moments \u222b x\u0304dm and \u222b y\u0304 dm about axes through the center of mass are zero (cf. Volume 1, Section 4.3). Therefore, using \u0398c = \u222b r\u03042dm = \u222b (x\u03042 + y\u03042) dm, x2 c + y2 c = r2 c , m = \u222b dm the parallel-axis theorem is obtained: \u0398a = \u0398c + r2 c m . (3.20) From (3.18) we find the relation r2 ga = r2 gc + r2 c for the radii of gyration. a b As an illustrative example we consider a slender homogeneous rod (mass m), see Fig. 3.17a. With dm/m = dr/l we obtain the moment of inertia \u0398A = \u222b r2 dm = m l l\u222b 0 r2 dr = ml2 3 (3.21a) 3.2 Kinetics of the Rotation about a Fixed Axis 151 with respect to an axis that is perpendicular to the rod and passes through point A. Here, we replaced the subscript a (reference axis) by the subscript A (reference point). In the following we will use both notations. If we choose a reference axis that passes through the center C of the rod, (3.20) leads to \u0398C = \u0398A \u2212 ( l 2 )2 m = ml2 12 . (3.21b) We now determine the moment of inertia of a homogeneous circular disk with mass m, thickness t, and radius R. We choose the reference axis c-c that is perpendicular to the plane of the disk and passes through its center (Fig. 3.17b). With the area element dA = 2 \u03c0r dr we obtain \u0398c = \u222b r2 dm = \u03c1t \u222b r2 dA = 2 \u03c0\u03c1t R\u222b 0 r3 dr = \u03c0 2 \u03c1tR4 = mR2 2 . (3.22) The moment of inertia \u0398c depends on the mass m and the radius R; it is independent of the thickness t. Therefore, the result (3.22) is also valid for a circular cylinder of arbitrary length. E3.5Example 3.5 Determine the moment of inertia of a homogeneous solid sphere (mass m, radius R) with respect to an axis that passes through the center C. Fig. 3.18 c C r R c dz z Solution We consider the sphere as being composed of circular disks with infinitesimal thickness dz (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003916_s0005117913080079-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003916_s0005117913080079-Figure3-1.png",
        "caption": "Fig. 3.",
        "texts": [
            " However, it is clear that impact motions of the mechanism are possible if only one of the pistons interacts with the anvil. Then the surface S will be a smooth cylindrical surface with equation either x = f1(\u03c4) or x = f2(\u03c4). Using equations x = f1(\u03c4), x = f2(\u03c4), we can get a relation on the mechanism parameters (\u03b5\u2212\u0394k) = \u03bc \u221a (1\u2212 \u03b3 cos\u03d5)2 + \u03b32 sin2 \u03d5, (3) which distinguishes, in the space of parameters, regions where motions with impacts by either both percussion pistons or one of them exist. These regions are shown on Fig. 3 in the plane of generalized parameters ( \u03b3, (\u03b5\u2212\u0394k \u03bc )2) for different values of the stationary shift in phase eccentricities \u03d5. Motion modes of the first kind are possible in the mechanism for parameter values that belong to the dashed regions; modes of the second kind, in the rest of the space. Obviously, the location of the corresponding regions shown on Fig. 3 may significantly influence preliminary tuning of the mechanism for a certain operational mode. AUTOMATION AND REMOTE CONTROL Vol. 74 No. 8 2013 Thus, the description of the phase space structure and the behavior of phase trajectories in it imply that as a section plane [3] we can take the surface S and study the dynamics with the method of point maps [3, 4]. In this case, it means mappings of the section plane S onto itself. 4. CONSTRUCTING A POINT MAP Let M0(\u03c4 = \u03c40, x = f1(\u03c40), x\u0307 = x\u03070) \u2208 S1, M1(\u03c4 = \u03c41, x = f2(\u03c41), x\u0307 = x\u03071) \u2208 S2, M2(\u03c4 = \u03c42, x = f1(\u03c42), x\u0307 = x\u03072) \u2208 S1 be three consecutive points that belong to the surface S; then a transformation T = T2 T1 of the points M0 T1\u2212\u2192 M1 T2\u2212\u2192 M2 can be written as T1 \u23a7 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u0394k \u2212 \u03bc\u03b3 cos(\u03c41 \u2212 \u03d5) = \u2212\u03bc\u03bb1(cos \u03c41 \u2212 cos \u03c40) \u2212 (\u03c41 \u2212 \u03c40)(\u03bc\u03bb1 sin \u03c40 + \u03bc\u03b3\u03bb2 sin(\u03c40 \u2212 \u03d5)) \u2212\u03bc\u03b3\u03bb2(cos(\u03c41 \u2212 \u03d5)\u2212 cos(\u03c40 \u2212 \u03d5)) \u2212 p (\u03c41 \u2212 \u03c40) 2 2 + x\u03070(\u03c41 \u2212 \u03c40) + \u03b5\u2212 \u03bc cos \u03c40; x\u03071 = \u2212R(\u03bc\u03bb1(sin \u03c41 \u2212 sin \u03c40) + \u03bc\u03b3\u03bb2(sin(\u03c41 \u2212 \u03d5)\u2212 sin(\u03c40 \u2212 \u03d5)) \u2212 p(\u03c41 \u2212 \u03c40) + x\u03070 + (1 +R)\u03bc\u03b3 sin(\u03c41 \u2212 \u03d5); (4) T2 \u23a7 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03b5\u2212 \u03bc cos \u03c42 = \u2212\u03bc\u03bb1(cos \u03c42 \u2212 cos \u03c41) \u2212 (\u03c42 \u2212 \u03c41)(\u03bc\u03bb1 sin \u03c41 + \u03bc\u03b3\u03bb2 sin(\u03c41 \u2212 \u03d5)) \u2212\u03bc\u03b3\u03bb2(cos(\u03c42 \u2212 \u03d5)\u2212 cos(\u03c41 \u2212 \u03d5)) \u2212 p (\u03c42 \u2212 \u03c41) 2 2 + x\u03071(\u03c42 \u2212 \u03c41) + \u0394k \u2212 \u03bc\u03b3 cos(\u03c41 \u2212 \u03d5); x\u03072 = \u2212R(\u03bc\u03bb1(sin \u03c42 \u2212 sin \u03c41) + \u03bc\u03b3\u03bb2(sin(\u03c42 \u2212 \u03d5)\u2212 sin(\u03c41 \u2212 \u03d5)) \u2212 p(\u03c42 \u2212 \u03c41) + x\u03071) + (1 +R)\u03bc sin \u03c42"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002982_amr.317-319.281-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002982_amr.317-319.281-Figure2-1.png",
        "caption": "Fig. 2. Deformation of the flexspline under load",
        "texts": [
            " However, to work properly, various bearings (including flexible bearing) need to have certain gap. In the process of wave deforming, flexspline will move periodically in the circumferential and axial direction on the mate surface between the bearing outer ring and the flexspline. Ivanovo pointed out in [5] that while the flexspline is under the load, the flexspline gear ring cannot maintain the shape under no-load. In other words, the mesh arc of the flexspline is separated from the wave generator. Depicted in Fig.2, circle 1 stands for the nominal profile of the flexspline contacting the wave generator without any gap or contact deformation, circle 2 stands for the wave generator profile with a gap \u03b4uniformly distributed along the whole profile, and circle3 (dashed) stands for the flexspline profile under loads. Under certain loads, some areas of the flexspline are closer to the wave generator and thus deformed, while other areas do not contact the wave generator and yield distortion. When the wave generator rotates in clockwise direction, the points on arc AB\u2032and A B\u2032 cannot deform freely in the radial direction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.102-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.102-1.png",
        "caption": "Fig. 2.102 4RPaRRR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology R\\Pa\\kR||R\\R",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.42-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.42-1.png",
        "caption": "Figure 12.42 A link with a spherical joint.",
        "texts": [
            "643) If we express the position vectors of ni and mi by 0ni = 0di\u22121 \u2212 0ri (12.644) 0mi = 0di \u2212 0ri (12.645) 0 i\u22121 di = 0mi \u2212 0ni (12.646) then Equation (12.643) will be the same as Equation (12.639). There are 2n vectorial equations of motion for an n-link multibody. However, there are 2(n + 1) forces and moments involved. Therefore, one set of force systems (usually Fn and Mn) must be specified to solve the equations and find the joints\u2019 force and moment. Example 757 A Link with Spherical Joint Figure 12.42 illustrates a link attached to the ground by a spherical joint at O. The free-body diagram of the link is made of an external force GFe and moment GMe at the end point, gravity mg, and driving force GF0 and moment GM0 at the joint. The Newton\u2013Euler equations for the link are GF0 + GFe + mg K\u0302 = m GaC (12.647) GM0 + GMe + Gn \u00d7 GF0 + Gm \u00d7 GFe = GI G\u03b1B (12.648) Example 758 Turning Arms Let us consider the turning uniform arm shown in Figure 12.43(a). Figure 12.43(b) illustrates the free-body diagram of the arm and its relative position vectors m and n: 0m = \u23a1 \u23a2\u23a2 \u23a2\u23a2 \u23a3 l 2 cos \u03b8 l 2 sin \u03b8 0 \u23a4 \u23a5\u23a5 \u23a5\u23a5 \u23a6 0n = \u23a1 \u23a2\u23a2 \u23a2\u23a2 \u23a3 \u2212 l 2 cos \u03b8 \u2212 l 2 sin \u03b8 0 \u23a4 \u23a5\u23a5 \u23a5\u23a5 \u23a6 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001444_j.elecom.2006.04.010-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001444_j.elecom.2006.04.010-Figure5-1.png",
        "caption": "Fig. 5. Cyclic voltammograms recorded on a poly(Mu)-coated platinum electrode in acetonitrile monomer-free solution containing 0.1 M Bu4NPF6 (scan rate, 100 mV s 1). Working electrode consists in platinum disc (/ = 1 mm) modified by oxidation in a 10 2 M Mu solution (5 sweeps of potential at 100 mV s 1 between 0.3 and +0.9 V) in CH3CN containing 0.1 M Bu4NPF6.",
        "texts": [
            "5 V, implying that poly(cyclopentadithiophene) matrixes undergo no anodic degradation. Furthermore, the p-doping level of poly(Mu), dp;Mu , is found to be higher than that of poly(Ms) whatever the potential, in accordance with the methylidene bridge contribution to the p-doping process as already mentioned by Zotti [20]. However, the more noticeable change of electronic properties occurs in cathodic investigations. CV\u2019s recorded in the negative potential range on a platinum electrode modified with a poly(Ms) film (Fig. 4) or a poly(Mu) film (Fig. 5), show, in each case, a quasi-irreversible system (noted AredjAox), a reversible system (noted BredjBox) and an irreversible other one (noted Cred). To facilitate the understanding, electrochemical systems attributed to poly(Mu) film are written by adding an apostrophe. For poly(Ms) film, the quasi-irreversible system A appearing before the n-doping process only occurs on thin films of polymer and could be attributed to the reduction or abstraction of acidic cyclopentadienic protons leading to polyanion as already evoked by Zotti concerning the general electrochemical behaviour of poly[(RH)CPDT]\u2019s [21]",
            " The reversibility of system B is consistent with the n-doping process of the poly(Ms) matrix as attested by the charge exchanged corresponding to the transfer of a fractional charge per monomeric unit, accompanied by a mass increase in the forward (reductive) scan of potential and to a mass decrease in the backward (oxidative) scan when the working electrode is an AT-cut 9 MHz platinum-coated quartz crystal oscillator. The irreversible system C appearing only at the first potential sweep is consistent with a bielectronic cleavage of the SN bond taking into account previous studies [18]. Contrary to poly(Ms), poly(Mu) shows a fundamentally different cathodic behaviour taking into account the presence of an unsaturated organic spacer. As may be seen in Fig. 5, the irreversible system C 0, which appears at the beginning of the n-doping, seems to have moved towards the positive potentials compared to poly(Ms). It must be noted that after having observed the disappearance of the cathodic wave C 0, this irreversible system is not restored even after positive polarization of the electrode, implying that the first cathodic wave corresponds to the SN bond cleavage and not to a remnant peak as frequently observed for CP\u2019s. In recurrent sweeps, the main irreversible cathodic wave C 0 totally disappears after the first sweep and a cathodic current increases slowly from 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001756_9781119970422.ch6-Figure6.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001756_9781119970422.ch6-Figure6.3-1.png",
        "caption": "Figure 6.3 Scenario classification (All sensor nodes (x) are covered with the transmission ranges that depicted as (o) for some examples). (a) Structured scenario with a known spatial deployment of nodes and a controllable assist node. (b) Unstructured scenario with unknown node positions and unpredictable movement node.",
        "texts": [
            " For scenarios with controllable node mobility we design an integrated path planning algorithm. For all scenarios we design appropriate algorithms to collect energy information. Scenario Classification In the gMAP approach, we focus on three important types of scenarios that provide basic features to build more detailed realistic scenarios. 1. In a structured scenario we assume that the spatial deployment of sensor nodes is known a priori and that the mobility of the assist node is controllable (Figure 6.3(a)). 2. In a semi-structured scenario with an a priori known (or reliably estimated) spatial deployment of sensor nodes we assume that the mobility of the assist node to be controllable (Figure 6.3(a)). 3. In an unstructured scenario the topology is unknown (e.g. random spatial deployment) and the mobility of the assist node is assumed to be unpredictable and uncontrollable (Figure 6.3(b)). Our main driver for the scenario selection is the proof of the concept in extreme scenarios. Furthermore, in a realistic scenario, the spatial deployment of sensor nodes can be structured or known only partially. The mobility of the assist node can be either controllable or uncontrollable and may follow varied patterns. Path Planning of Assist Nodes Path planning is required for structured and semi-structured scenarios. The assist node plans its movement according to the positions of sensor nodes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001634_b978-0-12-374920-8.00426-4-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001634_b978-0-12-374920-8.00426-4-Figure1-1.png",
        "caption": "Figure 1 Cross section of Chlamydomonas axoneme showing the typical 9\u00fe 2 pattern. The diameter is B0.2 mm.",
        "texts": [
            " The organism of choice in this case is Chlamydomonas, a biflagellate green alga that allows easy isolation of flagella, use of classical genetics, and transformation with cloned genes.8 Importantly, the properties of axonemes or axonemal proteins found in this organism are in most cases shared by other kinds of cilia and flagella. The variety, composition, arrangement, and motile properties of flagellar dyneins are by far the most extensively studied using this organism. Thus, this chapter mainly deals with the current knowledge of the axoneme and axonemal dyneins obtained from Chlamydomonas. The internal core of cilia and flagella is called the axoneme (Figure 1). It is the structure that remains after detergent treatment of isolated cilia/flagella. When suspended in appropriate solutions containing ATP, axonemes can display beating like the cilia/flagella in live cells. Thus, the mechanism that produces beating is solely contained in the axoneme. As stated previously, most axonemes have the 9\u00fe 2 form (Figure 1), although some minor types are known to have a 9\u00fe0 structure. In very rare examples, axonemes are found that have 3\u00fe0,9 6\u00fe 0,10 12\u00fe0, 14\u00fe 0, or other patterns.11 Cilia and flagella without the central pair of microtubules, found in nature, usually display three-dimensional movements. The microtubules circularly arranged on the outer side of axoneme are called the outer doublet microtubules or outer doublets because they are composed of an ordinary cylindrical microtubule (the A-tubule) and an incomplete tubule (the B-tubule) fused to the sidewall of the A-tubule"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002003_978-3-642-27329-2_25-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002003_978-3-642-27329-2_25-Figure2-1.png",
        "caption": "Fig. 2. Sketch of the moving mass system",
        "texts": [
            " (6) So the anti-rolling system is virtually increasing the inertia moment (From I up to I C), damping (From 2N up to 2N B) and initial stability (From Dh up to Dh+A). Table 1 shows the effects using different control strategies. Fig. 1 shows the roll RAO curves using different control strategies, where the horizontal axis is frequency of wave input and the vertical axis is the roll RAO of ship. The control target of anti-rolling device is to keep the ship stable by viewing the wave moment as noise and reducing the noise: ( ) ( ) ( )2xx mI C N B Dh A Dh\u03c6 \u03c6 \u03c6 \u03b1\u2032 + + + + + = Fig. 2 shows the 3d model of the moving mass system. In this system[5], the rail of the moving mass is placed across the deck, and the mass is moving along the transverse direction of ship. The distance between the moving mass and ship gravity center is Zm. So the 3d model can be simplified as 2d model just considering the single-degree-of-freedom roll. Fig. 3 shows the 2d force analysis of the moving mass system. is the position of the mass; F is the force of actuator exerting on the mass; Mc is the control moment; is roll angle; a is the acceleration of the moving mass"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002907_s1052618811030022-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002907_s1052618811030022-Figure1-1.png",
        "caption": "Fig. 1. The schemes of planetary gears: (a) a single row train; (b) a double row train; (c) a double row train with two external toothings; and (d) a double row train with two internal toothings.",
        "texts": [
            "95 Planetary drives have found a wide utility in different engines, apparatuses, and devices, not only for change of rotary speeds with a given or variable contact ratio (depending on a transmission capacity), but for reproduction of a given trajectory (path generating mechanisms). Such possibilities result from the combination of rotary and sometimes progressive motions. Figure 1 shows the schemes of the simplest single row and double row planetary trains. Each of the considered schemes can be a differential with a number of degrees of freedom more than unity. Under additional impositions, the train may be transformed into a planetary gearbox (for an increase of contact ratio) or multiplicator (for an increase of angular speed supplied on a driven shaft with decreasing torque) [1]. In the general case, a shaft is recognized as the drive shaft under coincidence of rotary speed and force moment directions",
            " It is common for a contact ratio to use the Villis\u2019s method, when the angular speed, equal in value and opposite in direction to the angular speed of the cage (reversed train), is supplied to all links. As this takes place, the planetary train turns into one with fixed axes, and corresponding relationships for single row and double row gearboxes can be used. If the sun wheel, ring wheel, and cage are denoted by subscripts 1, 4, and H, respectively, and the fixed link is represented by the superscript, then, for the reversed single row train (Fig. 1a), we find [1] (1) Equation (1) is referred to as the basic equation of the three link differential. The relationship for a train with a fixed ring wheel (\u03c94 = 0) follows from Eq. (1) = \u03c91/\u03c9H = 1 \u23af . The equation in this form is acceptable for double row trains as well (Figs. 1b\u20131d). The relationship can be generalized on the con tact ratio from any planetary drive \u201ci\u201d to the cage \u201cH\u201d at fixed support wheel \u201cj\u201d as follows: = \u03c9i/\u03c9H = 1 \u2013 . For the scheme with one external toothing and one internal toothing (Fig. 1b) we find (2) The disadvantage of the Villis\u2019s converse method is a determined formalism; the basic kinematic char acteristic of the train is found with the analysis of motion and kinematic connections missing. Moreover, it is only suitable for the determination of contact ratios, when one of the links is fixed. Positions, speeds, and accelerations of any particles of the train at any one time can be defined accord ing to the superposition method and Lagrange motion equations [2]. This method is convenient upon complex analysis of the drive with certain energy and power characteristics [3, 4]",
            " For the point A at \u03c9 = \u03c91, yA = r1 we find (5) Likewise, for the cage axis C (\u03c9 = \u03c9H, yC = r1 + r2) we have (6) The angular speeds \u03c91 and \u03c9H are not the algebraic val ues. The sign \u03c91 on the drive link can be prescribed arbi trarily (plus, it is under anticlock wise rotation as viewed from the side of positive direction of axis z). The sign \u03c9H should thereafter be defined taking into account the kine matic connections on the contact of the planetary roller with the drive 1 (sun) wheel and the support 4 wheel. For the double row train with one external toothing and one internal toothing (Fig. 1b) from the condition of \u03c5x( )A \u03c91r1.\u2013= \u03c5x( )C \u03c9H r1 r2+( ).\u2013= speed equality in the point B on the contact between planetary roller 3 and fixed wheel 4 (yB \u2013 yC = r3) in terms of Eq. (6) we find the angular speed of planetary rollers (wheels 2 and 3 rotates with the same speed) (7) The negative sign shows that the directions of the rotation of the planetary rollers and the cage do not coincide. The result can be used for determination of the point A on the contact between the sun wheel and planetary roller 2 (yA \u2013 yC = \u2013r2) (8) By making the second members of Eq. (5) equal to second members of Eq. (8), we find the relationship between the angular speed of the sun wheel and the cage Hence, it follows from the relationship (2) in view of r1 + r2 = r4 \u2013 r3 that (9) or At z2 = z3, we find the relationship for the single row train (Fig. 1a) = \u03c91/\u03c9H = 1 + (z4/z1). The trains of these schemes have a wide application in multiple planetary roller ratios of average and large power at = 3, \u2026, 15 and efficiency = 0.98, \u2026, 0.96 [1]. The availability of several planetary rollers \u03c5x( )B \u03c5x( )C \u03c923r3\u2013 \u03c9H r1 r2+( )\u2013 \u03c923r3\u2013 0= = = \u03c923 \u03c9H r1 r2+( )/r3.\u2013= \u03c5A \u03c5C \u03c923r2+ \u03c9H r1 r2+( )\u2013 \u03c923r2.+= = \u03c5A \u03c91r1\u2013 \u03c9H r1 r2+( )\u2013 \u03c9H r1 r2+( ) r2 r3 \u2013 \u03c9H r1 r2+( ) r3 r2+ r3 .\u2013= = = i1H 4 \u03c91 \u03c9H r1 r2+( ) r2 r3+ r1r3 r1r2 r1r3 r2r2 r2r3+ + + r1r3 = = = = 1 r2 r1 r2 r3+ +( ) r1r3 + 1 r2r4 r1r3 += i1H 4 \u03c91 \u03c9H 1 r2r4 r1r3 + 1 z2z4 z1z3 ",
            "+= = = i1H 4 i1H 4 198 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 40 No. 3 2011 ALYUSHIN et al. allows decreasing the dimensions, improving the dynamics (the equilibration of the cage and the overall train and the discharge of bearing), and decreasing the weight in comparison with other types of drives under the same contact ratios and powers. The single row train at = 3, \u2026, 8 is additionally of a small axial size and widely used in aircrafts, remote control equipment, and so on [1]. For the drive with two external toothings (Fig. 1c; r1 + r2 = r4 + r3), from the condition on fixed wheel 4 (10) we find the angular speed of planetary rollers \u03c923 = \u03c9H(r1 + r2)/r3. Unlike the scheme on Fig. 1b the planetary rollers and the cage have an identical sense of rotation (the signs of angular speeds coincide). By analogy with Eq. (10) for the point A of the contact between the sun wheel and cage 2, we can write and, in terms of the result (10), we find Hence, (11) For the drive with two internal toothings (Fig. 1d; r1 \u2013 r2 = r4 \u2013 r3) from the condition in the point B it follows that \u03c923 = \u2013\u03c9H(r1 \u2013 r2)/r3. Along with Eq. (5), the speed of the point A can be written as (12) Solving the system of equations (5) and (12), we find or (13) Equations (11) and (13), as well as (9), can be written in terms of the number of teeth In the general case, the angular speeds of central wheels 1 and 4 may be considered known, and the speeds of planetary and cage rollers may be considered unknown. Kinematic links on the contact of teeth for the points A and B (r1 + r2 + r3 = r4) determine the system of equations (14) from solution of which we find Under \u03c94 = 0 the relationship between \u03c9H and \u03c923 is transformed to (7)",
            "\u2013= = \u03c5x( )B \u03c5x( )C \u03c923 y yP\u2013( )\u2013 \u03c9H r1 r2\u2013( )\u2013 \u03c923r3\u2013 \u03c9H r1 r2\u2013( )\u2013 \u03c923r3\u2013 0= = = = \u03c5x( )A \u03c5x( )C \u03c923 y yP\u2013( )\u2013 \u03c9H r1 r2\u2013( )\u2013 \u03c923r2\u2013 \u03c9H r1 r2\u2013( )\u2013 \u03c923r2.\u2013= = = \u03c91r1\u2013 \u03c9H r1 r2\u2013( )\u2013 \u03c9H r1 r2\u2013( )r2/r3+ \u03c9H r1 r2\u2013( ) 1 r2 r3 \u2013\u239d \u23a0 \u239b \u239e\u2013= = \u03c91/\u03c9H i1H 4 1 r2r4( )/ r1r3( ).\u2013= = \u03c91/\u03c9H i1H 4 1 z2z4( )/ z1z3( ).\u2013= = \u03c91r1 \u03c9H r1 r2+( ) \u03c923r2, \u03c94 r1 r2 r3+ +( )\u2013 \u03c9H r1 r2+( ) \u03c923r3,+= = \u03c923 \u03c94 r1 r2 r3+ +( ) \u03c91r1\u2013 r2 r3+ , \u03c9H \u03c91r1r3 \u03c94r2 r1 r2 r3+ +( )+ r1 r2+( ) r2 r3+( ) .= = JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 40 No. 3 2011 KINEMATIC RELATIONSHIPS IN PLANETARY DRIVES 199 For the drive with two external toothings (Fig. 1c) instead of (14), we find the following system from solution of which we find Similarly, for the drive with two internal toothings (Fig. 1d), we have As follows from Eq. (9), the cage and the sun wheel always have an identical sense of rotation under fixed support (ring) wheel 4, and the relationship between the speeds depends on the wheel diameters (the numbers of teeth). The directions of rotation can be both identical and opposite in double row drives with two external and two internal toothings. The drives when the second of the summands in equations (11) and (13) are close to unity and the angular speed of the cage is in significant excess of the angular speed of the sun wheel (possibly, by several orders) are of special interest. The driving gear in such trains is generally the cage, and they work as reduc tion gears. For example, the angular speed of the cage exceeds the angular speed of the sun wheel by over more than 50 times for a train with two external toothings (Fig. 1c), and the numbers of teeth z1 = z2 = 100, z3 = 101, z4 = 99 when retaining the center to center spacing (r1 + r2 = r4 + r3) and the same module of the teeth in all wheels. Under the numbers of teeth z1 = z3 = 100, z2 = 99, z4 = 101 (the condition of coaxiality is not satisfied, but the difference in center to center spacing is small and 2m), the contact ratio attains = = 10000 in terms of (11). The numbers of teeth need to satisfy not only the condition of coaxiality, but also the conditions of assembly, collocation of several planetary rollers on the common circle in plane, and the condition of the absence of gear locking. The drive with the numbers of teeth z1 = 54, z2 = 45, z3 = 44, z4 = 55 and the con tact ratio meet these conditions. The trains with two internal toothings (Fig. 1d) prevail at the expense of smaller dimensions and higher efficiencies. Under the number of teeth z1 = 200, z2 = 99, z3 = 100, z4 = 201 (at the same modules, the condition of coaxiality is met if z1 \u2013 z2 = z4 \u2013 z3), we discover = 1/(1 \u2013 99 \u00d7 201/100/200) = 198. It is usual to use such trains with one planetary roller in drives when there is a need to get a large contact ratio, i.e., in mechanical sirens, but their another application is unlikely in connection with low efficiency (<1%) [1]. The technique used can be spread to more complicated planetary drives, and the aforementioned equations allow determining the modes of operation resulting in the passage of the tool (a boring head) along the given trajectory with formation of flutes situated at regular intervals, which lead to the sponta neous rock failure and the increase of tunnel works at the decrease of expenditure of energy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002436_afrcon.2013.6757663-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002436_afrcon.2013.6757663-Figure1-1.png",
        "caption": "Fig. 1. Model of humanoid body using circular and elliptical capsules as bounding volumes for self-collision detection",
        "texts": [
            " Sugiura et al, in [5], used a collision avoidance method using null-space optimisation criteria and task intervals. Only one joint of the colliding segments was used to avoid self collisions. A virtual force was generated between the two colliding segments. Motions were continually blended between collision avoidance and target reaching. The priority between the two was changed depending on the distance between the segments. A self collision detection scheme for humanoids, based on elliptical capsules was developed in our previous paper [6]. (See Figure 1.) The self collision detection method was shown to be simple and quick, while providing a good fit to the humanoid form. This paper now extends the previous work by presenting a simple and effective self collision avoidance method for humanoid robots that uses the elliptical capsule self collision detection. This paper will begin with an overview of self collision detection using elliptical and circular capsules in Section II. Self collision avoidance is then formulated in Section III. A case study of human dance imitation is used to validate the self collision avoidance method in Section IV. 978-1-4673-5943-6/13/$31.00 c\u00a92013 IEEE An elliptical capsule as defined in [6], is as an elliptical cylinder capped by ellipsoids. A circular capsule is defined as a circular cylinder capped by spheres. As shown in Figure 1, the arms, legs and neck of the humanoid are modeled using circular capsules. The body is modeled using elliptical capsules. If the body has a waist joint, two elliptical capsules are used. The shoulder girdle is an ellipsoid and spheres are used to represent joints. The self collision detection method developed in our previous paper [6] is used to compute the collision distance between humanoid body segments and is briefly summarised here. For segments represented by two circular capsules, the critical points giving the shortest distance between the capsules axis are found using the common normal between the two axes",
            " Human dance imitation of a humanoid is used as a case study to test the collision avoidance developed. The motion capture of a human dance is used as the motion input. Due to the differences between the human and humanoid forms, motion capture data of humans has to be processed to fit the humanoid structures. Self collision free motion of the humanoid is then generated. Only collisions between the arms and the torso are investigated in the case study. A humanoid robot is simulated for use in the case study as shown in Figure 1. The humanoid\u2019s torso has a three DOF waist and two four DOF arms comprising of a three DOF glenohumeral joint and a one DOF elbow joint. The dimensions of the humanoid are shown in Table II: The human dance motion capture data used in this study, was captured by the CSIR and the University of Johannesburg. This data set has a large range of different upper body movements which other data sets tend to lack. The motion capture data gives the x, y, and z position coordinates of the wrist, elbow and glenohumeral joint for over 2100 frames of motion"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.28-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.28-1.png",
        "caption": "Fig. 2.28 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PRPPa (a) and 4PPRPa (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology P||R\\P||Pa (a) and P\\P||R\\Pa (b)",
        "texts": [
            " PM type Limb topology Connecting conditions 1. 4RRPaR (Fig. 2.26a) R||R\\Pa\\kR (Fig. 2.21a) The first two and the last revolute joints of the four limbs have parallel axes 2. 4RRRPa (Fig. 2.26b) R||R||R\\Pa (Fig. 2.21b) The three first revolute joints of the four limbs have parallel axes 3. 4PRPaR (Fig. 2.27a) P\\R\\Pa\\kR (Fig. 2.21c) The second and the last joints of the four limbs have parallel axes 4. 4PPPaR (Fig. 2.27b) P\\P\\kPa\\R (Fig. 2.21d) The last revolute joints of the four limbs have parallel axes 5. 4PRPPa (Fig. 2.28a) P||R\\P||Pa (Fig. 2.21e) The second joints of the four limbs have parallel axes 6. 4PPRPa (Fig. 2.28b) P\\P||R\\Pa (Fig. 2.21f) The third joints of the four limbs have parallel axes 7. 4PRPPa (Fig. 2.29a) P\\R||P\\Pa (Fig. 2.21g) Idem No. 5 8. 4PRPaP (Fig. 2.29b) P\\R\\Pa \\kP (Fig. 2.21h) Idem No. 5 9. 4PPPaR (Fig. 2.30a) P\\P\\kPa\\kR (Fig. 2.21i) Idem No. 4 10. 4PPPaR (Fig. 2.30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.106-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.106-1.png",
        "caption": "Fig. 2.106 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRPaR (a) and 4RRPaRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology R\\R||R\\Pa\\kR (a) and R\\R\\Pa\\kR||R (b)",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-FigureB-2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-FigureB-2-1.png",
        "caption": "Figure B-2. Output Angle per Tooth Hit on the Sun Gear",
        "texts": [
            " There are a total of N planet gears. The ring gear which is fixed to the ground meshes with planet gears, too. The carrier is the output. For planet gears (Figure B-1), they will get hit once when \u03b8C + (-\u03b8P) = 360\u25cb (B1) From Table 1, the following rotational relationship can be derived as: R S P S R S C P N N N N N N \u2212 = \u03b8 \u03b8 (B2) After substituting Equation (B2) into (B1), the output angle per tooth hit on the planet gear can be derived as: \u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b = R P C N N 1 360o\u03b8 (B3) For the sun gear (Figure B-2), it will get hit once when \u03b8S - \u03b8C = 360\u25cb/N (B4) From Table 1, the following rotational relationship can be derived as: R S R S C S N N N N 1+ = \u03b8 \u03b8 (B5) After substituting Equation (B5) into (B4), the output angle per tooth hit on the sun gear can be derived as: Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 10 Copyright \u00a9 2011 by ASME \u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b = R S C N N N o360\u03b8 (B6) For the ring gear (Figure B-3), it will get hit once when NC o360 =\u03b8 (B7) Table 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure11-1.png",
        "caption": "Figure 11: Three detections in the prototype system",
        "texts": [
            " The TMP and Hohlraum manipulation stages will keep moving until the geometric boundaries of TMP and Hohlraum turn to be coincidence. Figure 8: Components of optical image system Figure 9: Geometric boundaries of the subassemblies can be traced real-time 4. RESULTS According to the design of precision robotic assembly system, the prototype system is developed. In the prototype system, the manual stages are used to substitute for the auto liner stages (Figure 10). There are three detections (shade outline detection, coaxial position detection and force / torque detection) in the prototype system (Figure 11). Source of parallel light Keyence laser sensor CCD Source of coaxial light Long-focus lenCatoptron mirror Auto liner stages Proc. of SPIE Vol. 8418 841819-4 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 04/21/2013 Terms of Use: http://spiedl.org/terms The first experiment which is carried out on Hohlraum inserted into TMP by the prototype system is successful, and the finished Half Hohlraum Component is shown in Figure 13. The friction force of Hohlraum inserting is 3 N which is greater than the theoretical value of 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure20-1.png",
        "caption": "Figure 20 \u02d8 Castor sensors",
        "texts": [
            " TIRE-2 (milled tread) 100/80-16''50P Type Tubeless Cornering stiffness coefficient - Fz=721 N 11,22 1/rad Cornering stiffness coefficient - Fz=972 N 11,62 1/rad Cornering stiffness coefficient - Fz=1289 11,18 1/rad Relaxation length 0,186 m Wheel weight 8,65 kg Wheel polar moment of inertia 0,313 kgm2 Wheel diametral moment of inertia 0,190 kgm2 TABLE III \u2212 WHEEL/TIRE\u22122 CHARACTERISTICS 6 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/75795/ on 04/09/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Figure 19 \u02d8 Cornering stiffness vs load Dynamic tests Vibration modes The velocity of the test rig belt can be assigned by means of the rig controlled electric motor; the castor was instrumented to detect the wheel rotation speed, the castor angular position around the steer axis by means of an angular sensor (AS in Fig. 20) and the castor lateral accelerations measured in correspondence of the wheel-axis (accelerometers AR and AL in Fig. 20). The two accelerometers, placed at the ends of the spindle, can detect (Fig. 21) the component, along the wheel axis, of : - the tangential and the centripetal accelerations due to the castor rigid rotation around the steer axis.; - the lateral accelerations due to the bending fork; - the tangential and the centripetal accelerations due to the fork torsion deformation around the torsion axis that can be considered intersecting the spindle in its middle point. Summing the signals of the two accelerometers the centripetal accelerations vanish (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002453_amr.812.107-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002453_amr.812.107-Figure2-1.png",
        "caption": "Fig. 2. Flexural modulus of composite in difference ESCR condition",
        "texts": [],
        "surrounding_texts": [
            "Flexural Strength. The flexural strength measures the force required to bend a beam under three point loading conditions. Figs. 1 and 2 show the results of flexural strength (MOR) and modulus of elasticity (MOE). Flexural modulus and flexural strength values gradually decreased with increasing exposure time of 4 month. The effect of water diffusion through the material is a matrix plasticization which progressively reduces the mechanical performance of the composite. Water presence causes also delamination at layers interface. Matrix plasticization reduces the material stiffness in the first stage of immersion while delamination at the layer interface occurs after longer water absorption times intensity [6]. The addition of liquid natural rubber in the sample had reduce the interface tension between the mixture and increased adhesion interface that enhance properties of the materials.The effects of water absorption due to water immersion, the soil burial, and the natural weather have changed the mechanical properties of kenaf fibre unsaturated polyester composites. These changes have been occurred maybe due to the change in the structures and the properties of the fibres, matrix phase, and the interface between them [7]. Fracture Toughness. The notch beam test in three-point bending is employed to measure the fracture toughness based on Kic values. From fig. 3, fracture toughness of all sample condition was decrease with increasing time exposed. For expose environment and buried in soil the values has slightly increase at days 90 due heavily raining season at this time. During this period, absorption of water at the sample surface help to increase flexibility without damaging the inner fiber of reinforcement. For distill and sea water exposure, specimens exposed in sea water showed lower values of degradation of fracture toughness as compare with distill water. The density of the medium has significant influence over the rate of diffusion of moisture into material. It depends on the constituent particles of the liquid; its osmotic pressure and number of voids in material [8]. Hardness. The durometer hardness test was used to measure the relative hardness of this material. The test method is based on the penetration of a specified indenter forced into the material under specified conditions. The hardness value for all board in each condition is about 80-85 shore D as shown in Fig. 4. The sample that had been exposed to environment has highest hardness result as compared with others condition. O-H from moisture react with free radicals formed by irradiation of LNR had helped to strengthen the surface structure of the sample. Reaction of oxygen with free radicals causes chain scission, and recrystallization occurs, increasing crystallinity, which is higher with aging in an air environment (rich in oxygen). An increase in crystallinity causes an increase in hardness and a consequent change in mechanical properties [9]. Infrared Spectroscopy. A series of Fourier transform infra red (FTIR) spectra were obtained to detect any chemical interactionbetween difference environment exposure to the sample. Fig. 5 shows the spectrum of IR in difference ESCR condition. There is no new peak observed. The \u2013OH peak for sea and distill water condition is slightly broader as compare with other condition. This shows that there is an increasing of \u2013OH group in the composite due to OH absorption from the water. A hydroxyl group was observed for all condition peaks from the absorption band 3500\u20133300 cm \u20131 accept for buried in soil condition. There is no -OH peak been observed. An aromatic functional group (C\u2013C stretch in ring) was observed from the absorption band 1600 and 1475 cm \u20131 and an alkanes group (CH3) from the absorption band 1376 cm \u20131 .It can be note that the aromatic functional group can also appear from matrix polymer [10]. It was identified with existing stretching band on 1300 \u2013 1000 cm \u20131 (C\u2013O) .The carbonyl region (1800 \u2013 1600 cm \u20131 ) reveals probably the presence of carbonyl group from the isoprene chain of liquid natural rubber [10]. Scanning Electron Microscope (SEM). Scanning electron microscopic (SEM) of the composites are taken to study the fibre matrix interaction and fracture behavior. In Fig.6 (a), fiber fracture behavior mostly been observe in composite sample. In fig.6 (b), fiber and matrix were fractured together and it shown a good interfacial bonding between them. Figs. 6 (c) and (d) show that the failure mechanism of these composite was mainly by fiber pull out due to the weakness of interfacial strength between fiber and matrix. The absorption of water in the fiber might been alleged to be responsible for insufficient adhesion between fiber surfaces and matrix binders thus causes weak bonding between them. Fiber fracture dissipates less energy compared to fiber pull out [11]."
        ]
    },
    {
        "image_filename": "designv11_100_0003445_ht2012-58139-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003445_ht2012-58139-Figure5-1.png",
        "caption": "Fig. 5: Structural analysis boundary conditions definition.",
        "texts": [
            " The bottom surface of the workpiece for the electric potential is electrically isolated and hence current density is set as zero. As for the remaining three side surfaces, as the electrical conductivity of the workpiece is very high and these surfaces being far enough from the weld pool, it is assumed that the electrical potential at these surfaces is zero. Boundary Conditions for Structural (Stress) Analysis: The boundary conditions used for the structural analysis are as follows, with reference to Fig. 5. Since the problem is symmetric about the weld line, stress analysis is also performed for only one-half of the workpiece, utilizing the symmetry boundary condition. Surface ABCD: This surface is the symmetry face and is defined as a frictionless face in the FE solver. Surface A\u2019B\u2019C\u2019D\u2019: This surface is taken as clamped, i.e. holding the two workpieces together. In the FE solver, this face is defined as fixed, i.e. x y 0 z \u03d1 \u03d1 \u03d1= = = . Surfaces ADD\u2019A\u2019, ABB\u2019A\u2019 and DCC\u2019D\u2019: These three faces have no constraints defined and are hence free to deform under the thermal load",
            " All deformation/stress results discussed here are after the temperature of the workpiece reaches ambient conditions. Figure 9 shows the contours of the total deformation along the weld line and the total workpiece distortion. The un-deformed workpiece is also superimposed in the figure. It is to be noted that to emphasize the modeled deformation of the structure after welding, the distortions are multiplied by a factor of 100. Figure 10 shows the total deformation (on a true scale) of the thin plate after welding. The welded thin plate undergoes distortion due to thermal gradients. With reference to Fig. 5, we find that surface A\u2019B\u2019C\u2019D\u2019 was clamped (fixed) and hence shows no deformation. Greater deformation is observed on moving closer to the welding line, with maximum deformation at the beginning and end of the welding path. This can be explained on the basis of end-effect. The plots for directional deformation are not shown here, but it is seen that the symmetry face ABCD shows no deformation in the x-direction. This confirms correct set up of the model in the FE solver, as the symmetry boundary condition is upheld"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003224_amr.565.171-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003224_amr.565.171-Figure1-1.png",
        "caption": "Fig. 1 Three-dimensional model and gird model of the oil film",
        "texts": [
            " In this paper, in order to build the grid model of the film, firstly, the finite element model of the film is built by using \"Gambit\", and then the model is divided into the grid model. The grid is divided into more than 670,000 hexahedral grid cells. The bearing has five imports of oil chamber, which are set as \"Pressure inlet\". The exits are the two end face of the bearing, which are set as \"Pressure exit\". The inside surface and the outside surface of the film are set as \"Wall boundary\". Three-dimensional model and grid model of the oil film are 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. (ID: 132.239.1.230, University of California, San Diego, La Jolla, USA-04/06/15,10:17:48) Selecting calculation model and setting parameters. In this paper, implicit stationary model is adopted to calculate, and the physical model is set as turbulent flow. Flowing medium is NO.2 main shaft oil, the parameters of which are: the density of lubricant 3810 /kg m\u03c1 = ; the specific heat capacity of lubricant 2000J/kg KpC = \u22c5 ; the thermal conductivity of lubricant 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002528_s0219455412500393-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002528_s0219455412500393-Figure3-1.png",
        "caption": "Fig. 3. Three-dimensional eight-node element: De\u00afnition.",
        "texts": [
            " The mechanical and kinematics description of the motions is summarized in Table 1 while the related deformative modes are shown in Fig. 2. In t. J. S tr . S ta b. D yn . 2 01 3. 13 . D ow nl oa de d fr om w w w .w or ld sc ie nt if ic .c om by M C M A ST E R U N IV E R SI T Y o n 03 /1 0/ 13 . F or p er so na l u se o nl y. 2.2. Three-dimensional eight-node element We refer now to a 2h , 2h , 2h eight-node element still centered in the origin O \u00bc \u00f00; 0; 0\u00de of the \u00f0 ; ; \u00de reference system (see Fig. 3) and to the classical trilinear interpolation u\u00f0 ; ; \u00de \u00bc a0 \u00fe a1 \u00fe a2 \u00fe a3 \u00fe a4 \u00fe a5 \u00fe a6 \u00fe a7 ; v\u00f0 ; ; \u00de \u00bc b0 \u00fe b1 \u00fe b2 \u00fe b3 \u00fe b4 \u00fe b5 \u00fe b6 \u00fe b7 ; w\u00f0 ; ; \u00de \u00bc c0 \u00fe c1 \u00fe c2 \u00fe c3 \u00fe c4 \u00fe c5 \u00fe c6 \u00fe c7 : \u00f02:5\u00de The following strain expressions are obtained: \" \u00bc u; \u00f0 ; ; \u00de \u00bc a1 \u00fe a4 \u00fe a5 \u00fe a7 ; \" \u00bc 1 2 \u00bdu; \u00f0 ; ; \u00de \u00fe v; \u00f0 ; ; \u00de \u00bc 1 2 \u00bd\u00f0a2 \u00fe b1\u00de \u00fe a4 \u00fe b4 \u00fe \u00f0a6 \u00fe b5\u00de \u00fe a7 \u00fe b7 ; \" \u00bc 1 2 \u00bdu; \u00f0 ; ; \u00de \u00fe w; \u00f0 ; ; \u00de \u00bc 1 2 \u00bd\u00f0a3 \u00fe c1\u00de \u00fe a5 \u00fe c6 \u00fe \u00f0c4 \u00fe a6\u00de \u00fe a7 \u00fe c7 ; \" \u00bc v; \u00f0 ; ; \u00de \u00bc b2 \u00fe b4 \u00fe b6 \u00fe b7 ; \" \u00bc 1 2 \u00bdv; \u00f0 ; ; \u00de \u00fe w; \u00f0 ; ; \u00de \u00bc 1 2 \u00bd\u00f0b3 \u00fe c2\u00de \u00fe b6 \u00fe c6 \u00fe \u00f0b5 \u00fe c4\u00de \u00fe b7 \u00fe c7 ; \" \u00bc w; \u00f0 ; ; \u00de \u00bc c3 \u00fe c5 \u00fe c6 \u00fe c7 : \u00f02:6\u00de 1250039-5 In t",
            " So, the following relation is valid: IE \u00bc 2E h : In the following we de\u00afne the invariants by basing our measures only on nodal distances. In Table 7, the de\u00afnitions of the invariants used and their dependence on the related deformative parameters are given. In particular, the already seen extensional, shearing and hourglass modes were considered. We extend, at present, the formulation obtained in the previous two-dimensional case to the three-dimensional eight-node element. The deformative parameters are the ones de\u00afned in Tables 2 6. The de\u00afnitions of the numbering of the nodes refer to Fig. 3. We summarize in Tables 8 12 the de\u00afnitions of the invariants and their dependence on the related deformative parameters. In particular, Tables 8, 9, 10, 11 Table 7. Deformations of the four-node element: Invariants and parameters expressions. IE \u00bc \u00f0D\u00f0n1;n2\u00de \u00fe D\u00f0n3;n4\u00de\u00de=2 2h E \u00bc IE =2h IE \u00bc \u00f0D\u00f0n1;n3\u00de \u00fe D\u00f0n2;n4\u00de\u00de=2 2h E \u00bc IE =2h IS \u00bc D\u00f0n1;n4\u00de D\u00f0n2;n3\u00de S \u00bc IS \u00f0\u00f0h \u00de2 \u00fe \u00f0h \u00de2\u00de1=2=8h h IH \u00bc D\u00f0n1;n2\u00de D\u00f0n3;n4\u00de H \u00bc IH =4h h IH \u00bc D\u00f0n1;n3\u00de D\u00f0n2;n4\u00de H \u00bc IH =4h h In t. J. S tr . S ta b. D yn . 2 01 3. 13 . D ow nl oa de d fr om w w w "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.13-1.png",
        "caption": "Fig. 2.13 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRPRR (a) and 4RPRRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R||P||R||R (a) and R\\P||R||R||R (b)",
        "texts": [
            " 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No. 13 18. 4RRPRR (Fig. 2.10b) R||R||P||R\\R (Fig. 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2.1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003803_amr.605-607.1453-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003803_amr.605-607.1453-Figure4-1.png",
        "caption": "Fig. 4 3D platform positioning accuracy experiment",
        "texts": [
            " Materials and methods Materials: AURORA space magnetic locator, ZK-3000 ultrasound scanner, Lenovo Think centre M8000t PC machines, Surgery navigation software, active plus passive ultrasound-guided robot for liver cancer coagulation therapy, Multi-modal puncture phantom (Model071, the CIRS, the United States) Methods: (1)Experiment 1: Make use of the navigation and positioning method that was mentioned in the article, calculated the transformation matrix T which is the transformation matrix of 3D platform space mapping to the magnetic locator transmitter space, and then select another four points whose coordinates in 3D platform space was known (as shown in figure 4, A, B, C and D). Then, with a probe to collect the coordinates of four points in the magnetic locator transmitter space, using the formula to calculate the coordinates of these points in move in 3D platform space, and compared the calculated value with actual value, verify the navigation and positioning accuracy of 3D platform. (2) Experiment 2: As shown in figure 5, make use of a target point in the multi-modal puncture phantom (Model071, the CIRS, USA) as the experimental puncture target. According to the information collected by ultrasonography acquisition subsystem, robot carried out the puncture in accordance with the planning puncture path and to see if can access the puncture point and target spot, to verify the navigation and positioning accuracy of this method; fixed this group points (the puncture point and target spot), repeat the above experiment, verify that the robot positioning repeatability"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002797_amr.468-471.2141-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002797_amr.468-471.2141-Figure4-1.png",
        "caption": "Fig. 4 Assembled structure and finite element model",
        "texts": [
            " And the beam element of the joints can be created by using the user-defined element Matrix 27. The element sizes of the mesh operations influence the characteristics of the joints. Consequently, this investigation modified the element sizes of the structure, obtained the corresponding stiffness and damping matrices of the joints and compared the modal frequency of the assembled structure to reduce the influence of the element sizes. In order to validate the proposed modeling method of the joints, the assembled structures have been meshed. As shown in Fig.4, the assembled structure is mainly composed of the three sections of cross-rail, the ram and the cargo chute box. The three sections of cross-rail are clamped by the bolts with high strength. And the hydrostatic guideway connect the joints between the cross-rail and the cargo chute box, the ram and the cargo chute box. The stiffness and damping value are vulnerable to the interfacial pressure, the material of joints, the size, excitation frequency, machining method of the joints, the types of the joints, etc"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001303_pamm.200610036-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001303_pamm.200610036-Figure2-1.png",
        "caption": "Fig. 2 The PRANDTL-REUSS material model, the friction regularization model",
        "texts": [
            " Furthermore the dependency of a friction force orientation on a small perturbation seems unrealistic from a macroscopic point of view. Here we want to outline the functioning of the proposed regularization method without concentrating on the choice of contact points. As pointed out above, the goal is to use smooth transitions of the frictional force orientations between sticking and sliding states in both directions. This is achieved by regularizing the sticking state with a visco-elasto-plastic element (cf. Figure 2). Figure 2 specifies the structure of the equations for time-independent plasticity, also called PRANDTL-REUSSequations [2] in continuum mechanics. In our case Figure 2 identifies real spring, damper, and plastic elements, but their deformation should be analogous to the rheological model. \u2217 Corresponding author: e-mail: wolfgang.stamm@luk.de, Phone: +49 7223 941 609, Fax: +49 7223 941 7436 \u2217\u2217 Corresponding author: e-mail: alexander.fidlin@luk.de, Phone: +49 7223 941 9040, Fax: +49 7223 941 7436 \u00a9 2006 WILEY-VCH Verlag GmbH & Co. KGaA, Weinheim \u00a9 2006 WILEY-VCH Verlag GmbH & Co. KGaA, Weinheim The displacement of a contact point x during sticking is divided into elastic displacement z and plastic displacement w: x = z + w (1) Due to the series connection, the forces in elastic and plastic element are equal: Ff = Fel = Fpl (2) The elastic force is a spring with some pre-stressing zp and a damper: Fel = \u2212c(z + zp) \u2212 dz\u0307 (3) Analogous to the material model in [2], the plastic deformation is defined as: \u2212\u03bbFpl = w\u0307 (4) where \u03bb controls the occurence of plastic deformation: \u03bb = { 0, \u2016Fpl\u2016 < \u00b5st \u2016w\u0307\u2016 /\u00b5dyn \u2016Fpl\u2016 \u2265 \u00b5st (5) with a maximum static friction force \u00b5st and a steady state friction force \u00b5dyn for gross sliding"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003009_imece2013-62977-Figure16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003009_imece2013-62977-Figure16-1.png",
        "caption": "Fig. 16: Catapult design drawing [9]",
        "texts": [
            " Elastic band concept: the elastic band would be at high risk of snapping and, as a result, it received the lowest score. Counterweight design: the counterweight could possibly become detached. However, the frequency of this occurring was deemed less than the possibility of the elastic band snapping. Downloaded From: https://proceedings.asmedigitalcollection.asme.org on 04/09/2019 Terms of Use: http://www.asme.org/about-asme/terms-of-use 9 Copyright \u00a9 2013 by ASME The final catapult design that continued to be fully evaluated and developed is shown in fig. 16. The parts, as identified in fig. 16, are as follows: 1. Short Display Side (2) 8. Trigger 2. Extension Spring 9. Pulley 3. Angle Support (2) 10. Long Display Side (2) 4. Brace (10) 11. Lever Arm 5. Trigger Spring 12. Scoop 6. Ratchet Wheel 13. Base 7. Support Structure In addition to the catapult, the MAE design team designed and built a game board onto which the catapult projectile can be aimed for certain designated point values. While a great deal of research and analysis was required to determine the best design concept, after the spring catapult design was selected even more detailed analysis was required in order to assure the highest quality and safest product"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002320_iscid.2013.80-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002320_iscid.2013.80-Figure1-1.png",
        "caption": "Figure 1. Diagram of coordinate system",
        "texts": [
            " Without considering the sea conditions of emission, the kinematics model and dynamics model of vehicle are established by analysis of the force working on vehicle moving underwater, based on the theorem of the motion of mass center and the theorem of momentum of the rotation around the center of mass. The motion in six degrees of freedom of the vehicle is determined. And then, the position and attitude of vehicle in the space at any instant are determined. II. THE ESTABLISHMENT OF COORDINATES According to the need of model of underwater motion, 3 right-handed coordinates is established as shown in Fig. 1, namely the launch coordinate, the body coordinate and the velocity coordinate. The launch coordinate: the coordinate origin o is corresponding to center of buoyancy of the vehicle; the axis ox is pointing to the emission direction; the axis oy is upward vertically. The body coordinate: the coordinate origin 1o is corresponding to center of buoyancy of the vehicle; the axis 11xo is pointing to the head of vehicle along the longitudinal axis; the axis 11 yo is vertically upward in the longitudinal plane of symmetry"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002941_kem.462-463.366-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002941_kem.462-463.366-Figure1-1.png",
        "caption": "Fig. 1 (a) Location of failed gear and (b) key dimensions (mm)",
        "texts": [
            " Two consecutive failures were reported by the maintenance section of the factory. The first, the reducer gearbox shaft broke after approximately one year in service. The new shaft was produced according to the original drawing to replace the failed shaft. A crack in the herringbone gear, the second failure, was reported only one day after the operation. The failed gear has 20 teeth and the face width of 650 mm. The reducer gearbox was driven directly by 1800 kW AC motor at the speed of 750 rpm. Relevant layout of the gearbox and key dimensions are shown in Fig. 1 (a) and (b). The failed herringbone gear was first inspected visually and macroscopically. Dye penetrant technique was employed to enhance visual inspection of the crack and to reveal the nature of the crack more clearly. The material in the vicinity of the crack was cut from the gear and metallographic specimens were prepared for optical microscopy examination, micro-hardness measurement and scanning electron microscope (SEM) examination. Chemical analysis of the failed gear material was performed in order to identify the type of steel used"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.19-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.19-1.png",
        "caption": "Figure 4.19 Lubrication with auxiliary gear",
        "texts": [
            " The lower the circumferential speed of the pinion, the greater the dive in an oil bath should be enabled. Low-speed gears of the second and third step can be immersed to a depth of up to half a kinematic radius of a gear. It is not recommended for a fast-running gear, nor for the gears of the second and third steps to be deeply submerged in oil. Gears on remote shafts of multi-step vertical gear trains and gears of fast-running steps of reducers often do not reach the oil level, thus they are lubricated by an auxiliary gear or a ring (Figure 4.19). An auxiliary gear is usually made of polymer or textolite. The width of this gear is commonly taken from 0.3 to 0.5 of the width of the basic gear. When the supply of lubricant in the roller bearing is aggravated due to specific design or if it is deficient (e.g. low peripheral speed, thus lubricant is insufficiently dispersed on the walls of the housing), the problem can be fixed by: Scrapers that remove lubricant from the side of the gear rim and put it in the bearing, Grooves in the rim of the upper part of the housing through which the lubricant flows in the bearing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003854_20120905-3-hr-2030.00044-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003854_20120905-3-hr-2030.00044-Figure1-1.png",
        "caption": "Fig. 1. Model of flat-footed biped robot with mechanical impedance at each ankle",
        "texts": [
            " To show effects of ankle elasticity achieving energy-efficient walking, we therefore analyze flat-footed passive dynamic walking with mechanical impedance at the ankles using Extended ROS (EROS) for walking analysis in the all walking 978-3-902823-11-3/12/$20.00 \u00a9 2012 IFAC 518 10.3182/20120905-3-HR-2030.00044 phases. Using EROS, we can show the equivalence between the walking of flat-footed robots and that of arc-footed robots. We finally discuss the effects of ankle elasticity on biped walking from the viewpoint of this equivalence. In this paper, we analyze the passive dynamic walking of the flat-footed biped robot as shown in Fig. 1. This model has a spring, a damper, and an inerter at each ankle. The dynamic equation for this model is given by M\u03b2(q)q\u0308+C(q, q\u0307)q\u0307+G(q) = \u2212Kmq\u2212Dmq\u0307+JT c \u03bb, (1) where q = [\u03b81, \u03b82, \u03b83, \u03b84, x1, z1] T is the position vector, \u03b8i is a link angle, (x1, z1) is a position of the stance heel, M\u03b2(q) \u2208 R 6\u00d76 is an inertia matrix that contains the moment of inertia due to ankle inerters, C(q, q\u0307) \u2208 R 6\u00d76 is a matrix of Coriolis and centrifugal force, G(q) \u2208 R 6 is a gravitational vector, Km \u2208 R 6\u00d76 is a spring matrix, Dm \u2208 R 6\u00d76 is a damping matrix, JT c \u2208 R N\u00d76 is a constraint Jacobi matrix, and \u03bb \u2208 R N is a vector of constraint force, N is a number of constraints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001093_012079-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001093_012079-Figure1-1.png",
        "caption": "Figure 1. Diagram of the impact on the plane of bars of different stiffness: absolutely rigid (a), flexible thread (b) and real bar (c); (d) \u2013 diagram of normal pressures of a real bar on the plane.",
        "texts": [
            " Our research shows that there are 3 points in the work of the bar: International science and technology conference \"Earth science\" IOP Conf. Series: Earth and Environmental Science 720 (2021) 012079 IOP Publishing doi:10.1088/1755-1315/720/1/012079  embedding the bar in the contaminant layer;  sliding the bar along the contact path;  the output of the contact. The behavior of the bar will be determined by both the rigidity of the bar itself and the hardness of the surface to be treated. In this regard, 3 cases can be considered: \"absolutely rigid rod\", \"absolutely flexible\" and the actual one (figure 1). An absolutely rigid rod (figure 1 (a)) affects the surface with its outer end. When working, the rod does not deviate from its original position. The action of such a bar is similar to the action of a cutter. a) b) c) d) Such a rod is not able to copy the relief of the root and is not able to remove dirt from the recesses. In addition, it will injure protruding parts on the root surface. The shear resistance acting on the outer end of a rigid bar is defined as: Tssf = S \u00b7 psp = psp \u00b7 d \u00b7 h, (1) where Tssf is the soil shear force, sdf; S is the cross-sectional area of the chip, cm 2 ; Psp is specific resistance to soil shear, sdf/cm 2 ; h is the depth of penetration of the rod, cm; d is the bar diameter, cm. The complete opposite is an absolutely flexible rod (Figure 1 (b)), which, without rigidity, is not able to develop a shear force. The vast majority of brush rods occupy an intermediate position, i.e. they are flexible rods with some elasticity (Figure 1 (b)). Successful separation of a particle due to friction forces is possible if: S \u00b7 psp \u2264 Ffr = \u0192 \u00b7 PN, (2) where psp is the specific resistance to soil shear, kgf/cm 2 ; S is the cross sectional area of the chip, cm 2 ; International science and technology conference \"Earth science\" IOP Conf. Series: Earth and Environmental Science 720 (2021) 012079 IOP Publishing doi:10.1088/1755-1315/720/1/012079 Ffr is friction force between the bar and the cleaning surface; PN is normal bar pressure on the surface to be treated; \u0192 is a coefficient of friction"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002414_amm.99-100.857-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002414_amm.99-100.857-Figure5-1.png",
        "caption": "Figure 5 Pressure distribution Figure 6: Velocity Vector",
        "texts": [
            " Different numbers of grid were tested to reach the grid independent. Table 2 shows the simulation results. Because of the pressure and stress is very less when the bearing moves in X direction, only the force of Y and Z direction are presented here. Comparing the numerical results of different grid number, we can get the conclusion that the number of grid has no influence to the result of numerical simulation when the number of grid reaches 998000. The comparison of different grid showed in table 2 Figure 5 and figure 6 are the numerical simulation result of 3D journal bearing respectively. The comparison between the result of bearing\u2019s numerical simulation and reference [14] shows that the load of bearing equals 30.2N using CFD, and this result shows good match with reference. It indicates that the numerical computation method this paper has adopted can analyse and research the correlative parameters of bearing, for example: the lubrication state of bearing and load and so on. This paper built a sold foundation for 3D numerical simulation analysis method to be used for optimal design for journal bearing and analysis and computation for lubrication state and load of bearing of complicated geometry-shape"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure10-1.png",
        "caption": "Figure 10 FDM bead width",
        "texts": [
            " Currently, the build temperature is automatically controlled via a microchip material canister sold by Stratasys. Despite this restriction, bead width, raster orientation and raster air gap remain available to be adjusted by the end-user on the 400MC. The bead width is dictated by the FDM user to provide a width to the toolpaths taken by the process. There are two types of bead width, contour and raster. The contour is the outside wall of an X\u2013Y sliced plane. The raster is the filled pattern inside the contour. Figure 10 illustrates the bead width cross section of a raster pattern. Raster angle is defined as the relative angle placement of contours in the Z-direction. If a cross section was taken of a tensile bar built in the Z\u2013direction, Figures 11\u201313 would highlight a 0\u00b0, 90\u00b0 and 45\u00b0 raster orientation pattern, respectively. The next parameter allowed for adjustment is perimeter to raster air gap. This feature is defined as the amount of bead overlap between the interior fill material and the outside contour of a planar cross section"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003808_appeec.2012.6307442-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003808_appeec.2012.6307442-Figure1-1.png",
        "caption": "Fig. 1 (a) Toroidal core (b) General dimensions",
        "texts": [
            " Gyimesi and Lavers have published a semi-analytical solution for a magnetic field excited by a long current through the center of a nonconducting toiopid of circular cross section[1]. This paper presents a formula to calculate iron loss in transformer windings on ferromagnetic cores. A comparison of the 978-1-4577-0547-2/12/$31.00 \u00a92012 IEEE calculated values with measured values is presented. II. BASIC IMPEDANCE FORMULA A. Calculation of Self and Mutual Impedances The work presented in[2] is wonderful and outstanding due to the analytical calculation of self and mutual impedance illustrated in Fig. 1. Two coils are wound concentrically on the iron core. The mutual impedance between the two coils is given by 12 a b c dZ Z Z Z Z= + + + (1) 0 1 2 max( , ) min( , )a r aZ j N N ra l r a \u03c0\u03c9\u03bc= (2) 0 1 2 1 1 1 4 ( max( , )) ( min( , )) cos( ) b n n n n Z j N N ra l I r a K r a z \u03c0\u03c9\u03bc \u03b2 \u03b2 \u03b2 \u221e = = \u00d7 \u2211 (3) 2 1 0 1 2 0 2 ( ) 1 ( ) z c I mbbZ j N N l mbI mb \u03bc\u03c0\u03c9\u03bc \u23a7 \u23ab = \u2212\u23a8 \u23ac \u23a9 \u23ad (4) 1 2 0 1 2 1 1 2 1 11 2 1 2 4 ( , ) ( , )d n n n n n N NZ j P r r P a a l h h w w \u03c0\u03c9\u03bc \u03b2 \u03b2 \u03b2 \u03b2 \u221e = = \u2211 1 1 1 2 1 1 ( ) ( )( , ) cos( ) ( ) n n n n n n I b F bQ w w Z K b \u03b2 \u03b2\u03b2 \u03b2 \u03b2 \u03b2 \u00d7 (5) In these equation, 0zm j\u03c9\u03bc \u03bc \u03c3= , 2 \u03c0n n l\u03b2 = (where l is the length of the magnetic circuit). General dimensions are defined in Fig 1 and 0I , 1I , 0K and 1K are modified Bessel functions. The functions 1P , 1Q and 1F are defined in Appendix. When self impedance are required (e.g. 11Z ), These are computed as mutual impedance between the coil and itself. This involves setting 0z = in Eq. (5) and 0.2235( )z h w= + in Eq. (3), where h is the height of the coil cross section and w is its width. B. Calculation of leakage reactance \uff08a\uff09 (b) Fig.2 Leakage inductance Leakage inductance is the property of one winding with 2\u03c3\u03c6 +- +- 2u 2i 12\u03c6 coil1 coil2 +- +- 1u 1i 1\u03c3\u03c6 21\u03c6 coil1 coil2 respect to another"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.1-1.png",
        "caption": "Figure 12.1 A body point mass moving with velocity GvP and acted on by force df.",
        "texts": [
            "2 Rigid-Body Dynamics Employing Newton and Euler equations of motion (2.58) and (2.59) for a particle, we develop the equations of motion of rigid bodies. We usually express the equations of motion of rigid bodies in a local Cartesian coordinate frame attached to their mass center. It is a more practical method. 12.1 RIGID-BODY ROTATIONAL CARTESIAN DYNAMICS Consider a rigid body B with a fixed point in a global coordinate frame G, as shown in Figure 12.1. The rotational equation of motion of the rigid body is the Euler equation GM = Gd dt GL (12.1) BM = Gd dt BL = B L\u0307 + B G\u03c9B \u00d7 BL = BI B G\u03c9\u0307B + B G\u03c9B \u00d7 (BI B G\u03c9B ) (12.2) where L is the angular momentum of the body, GL = GI G\u03c9B (12.3) BL = BI B G\u03c9B (12.4) and I is the mass moment matrix of the rigid body B, BI = \u23a1 \u23a3 Ixx Ixy Ixz Iyx Iyy Iyz Izx Izy Izz \u23a4 \u23a6 (12.5) GI = \u23a1 \u23a3 IXX IXY IXZ IYX IYY IYZ IZX IZY IZZ \u23a4 \u23a6 (12.6) The expanded forms of the B-expression of the Euler equation are Mx = Ixx \u03c9\u0307x + Ixy \u03c9\u0307y + Ixz \u03c9\u0307z \u2212 (Iyy \u2212 Izz ) \u03c9y\u03c9z \u2212 Iyz ( \u03c92 z \u2212 \u03c92 y )\u2212 \u03c9x ( \u03c9zIxy \u2212 \u03c9yIxz ) (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001302_ias.2006.256560-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001302_ias.2006.256560-Figure6-1.png",
        "caption": "Figure 6. photograph of the three rotors, skewed (left), unskewed (center) and unskewed with oper slots (right)",
        "texts": [
            " RESULTS The proposed space-vector state model of the induction machine including rotor slotting effects has been implemented in numerical simulation and verified experimentally on a properly devised test setup. With regard to the numerical simulation, the model has been implemented in the MatlabSimulink environment. With regard to the experimental verification of the model, three motor prototypes have been built with the same stator and different rotors. As recalled in \u00a7III. All the three motors have 2 pole pairs, 36 stator slots and 28 rotor slots. Particularly, motor 1 has a skewed rotor, motor 2 has an unskewed rotor and motor 3 has an unskewed rotor with open rotor slots. Fig. 6 shows the photograph of the three rotors. To assess the proposed model, two kinds of representation have been made. First, after supplying the motor both in numerical simulation and experimentally with a sinusoidal waveform at the rated frequency of 50 Hz and voltage of 220 V, in the first representation the steady-state stator current spectrum has been computed. In he second, the stator current space vector locus at steady state has been drawn (isQ vs isD). If supplied with a pure sinusoidal voltage, neglecting all the spatial harmonics in the machine, the stator current locus would be a circle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003082_icems.2013.6754391-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003082_icems.2013.6754391-Figure12-1.png",
        "caption": "Fig. 12. Distribution of magnetic flux density (6 poles\u21922 poles).",
        "texts": [],
        "surrounding_texts": [
            "Fig. 10 shows the core loss of the stator at no load. If the number of poles in the motor is reduced, the lower frequency in the motor leads to a decrease in its core loss. The motor can decrease the core loss in the high speed region by about 34.3%. Therefore, pole changing leads to high efficiency in the motor."
        ]
    },
    {
        "image_filename": "designv11_100_0003245_amr.433-440.7247-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003245_amr.433-440.7247-Figure8-1.png",
        "caption": "Figure 8. Control winding input current while speed sudden change",
        "texts": [
            "0pml . Control side: pole pairs of control winding 2pc = , stator resistance \u2126= 835.1csr , stator inductance \u0397= 12362.0csl , rotor resistance \u2126= 544.1crr , rotor inductance \u0397= 12362.0crl , mutual inductance \u0397= 14986.0cml . Uniform chance curve of machine given speed is shown in figure5; sudden chance curve of machine given speed is shown in figure6; control winding input current curve while speed uniform change is shown in figure7;control winding input current curve while speed sudden change in figure8. Cascade brushless double-fed machine is provided with the characteristics of dual-fed power supplying, achieve the structure of Non-contact , it is sized of potential applications in the medium pressure, high-capacity drive and the AC excitation variable speed constant frequency power field. But the complexity of structural characteristics cause difficulties to analysis, realization of mathematical modeling and computer simulation are the important ways to study of the motor. The key of cascade brushless double-fed machine modeling is to select a suitable two-phase coordinate system and determine the correct coordinate transformation relations so as to achieve the correct description of the relations between decoupling and electromagnetic, rotor speed two-phase coordinate system can meet the requirement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002864_sami.2013.6480983-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002864_sami.2013.6480983-Figure2-1.png",
        "caption": "Figure 2. Surface Constructor used for development and analysis of grindability of a new worm gearing [7,8]",
        "texts": [
            " The discussion of the contact situation, so the relative position of the \u03a6-parallel motion path to the surface F1 reveals the possible local undercut cases. Moreover the Reaching Model theory can handle the dangerous interferences of F1 and F2 surfaces as global cuts, global happenings in the R=R(\u03a6) functions. For detailed explanation see [1]. Using the simple but robust and general theory SC produces F2 as a grid of the Pk\u2019 points and gives a handy tool for detection of local undercuts and global cut situations, please refer to [2,3]. To give an example here, Fig. 2 shows the screenshot of the development of a new type worm gearing having changing pitch along the worm axis and presents the analysis of the grindability of the worm. Due to the central role of motion paths in the Reaching Model, SC can visualize them. Speed vectors and acceleration vectors also can be shown connected to these paths as presented in Fig. 3. It is well known from the kinematical projective geometry that the space of relative velocity and the special loci of this space, like the axis of the instantaneous screw axis determine the quality of contact"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002824_vppc.2012.6422690-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002824_vppc.2012.6422690-Figure1-1.png",
        "caption": "Fig. 1. Deviation of the wheel diameters",
        "texts": [
            " Deviation of the wheel diameters Even though the wheel diameter is managed within certain constraints, deviation is occurred all the time by spinning with no traction, sliding, friction, adhesion, and so on[J]. Once the deviation is occurred, it is getting worse and period of the wheel turning is getting shorten. In general, the deviation of wheel diameters is managed within 6[mm] from 880[mm] of wheel to prevent torque unbalance but detailed characteristics should be taken into account. B. Deviation of the wheel diameters Figure 1 shows two different sizes of wheels (Here, 0< x <I). In case the diameter of wheel 'A' is bigger than that of wheel 'B', the speed of the motor 'A' which is connected to wheel 'A' is lower than that of the motor 'B' which is connected to wheel 'B'. Because the speed of motor 'B' is bigger, the induced voltage of the motor 'B' is bigger than that of the motor 'A'. From a common VSI, the current of motor 'B' is smaller than that of motor 'A' according to the voltage equation. This bigger current of motor 'A' produces bigger torque and the wheel 'A' gains traction force mainly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.17-1.png",
        "caption": "Figure 2.17. Power-angle characteristics for salient pole generators.",
        "texts": [
            "65c), (2.65d) in (2.64a) and (2.64b), we obtain the expressions of active and reactive powers: P \u00bc EfV Xd sin di \u00fe V2 2 Xd Xq XdXq sin 2di (2.66) Q \u00bc Ef V Xd cos di cos2 di Xd \u00fe sin2 di Xq V2 (2.67) Considering the case of a round rotor generator, for which Xd \u00bc Xq \u00bc Xs, the expressions of the active and reactive powers from (2.66) and (2.67) reduce to P \u00bc EfV Xs sin di (2.68a) Q \u00bc Ef V Xs cos di V2 Xs (2.68b) As shown in equation (2.66), the active power characteristic of a salient pole generator (Figure 2.17a) is composed of a main sinusoidal component, overlapped by a second component of double frequency sinusoidal form. The magnitude of the second component is proportional to the difference between the magnitudes of the d- and q-axis components of the synchronous reactance and does not depend on the emf Ef. The second-order harmonic displaces the maximum of the nonsalient pole generator power characteristic; therefore, the critical angle dcr for which the maximum power is obtained is smaller than 90 ",
            " On the contrary, the power characteristic magnitude Pe increases for the same values Ef and Xd compared to the salient pole machine characteristic. Note that the increase of power Pe is important only for very small values of the emf Ef, when the magnitude of the fundamental power characteristic \u00f0EfV=Xd\u00de has the same order as the magnitude of the second harmonic \u00f0\u00f0V2=2\u00de\u00f0Xd Xq\u00de=\u00f0XdXq\u00de\u00de. Usually, when the emf Ef has a considerable large value, the magnitude of the second harmonic is less than 10\u201315% of the fundamental magnitude and, consequently, its influence on the power characteristic magnitude is very small (Figure 2.17b). Observation Since the representation of salient pole generators (for which Xd 6\u00bc Xq) requires large computational effort, in static stability analysis the saliency is omitted, and their equivalent reactance is taken equal to the direct synchronous reactance Xd. This assumption allows us to neglect the second-order harmonic of the power characteristic. A salient pole machine is represented as a round rotor machine even when, for accuracy, the effect of saliency needs to be taken into account. A fictitious emf is introduced in the round rotor generator model, which changes in terms of the generator operation, so that the active and reactive powers of the equivalent generator reflect with high accuracy the operation of the real salient pole generator"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002914_j.proeng.2012.09.161-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002914_j.proeng.2012.09.161-Figure1-1.png",
        "caption": "Fig. 1. Fabrication result of the PDMS microfluidic three-electrode cell chip",
        "texts": [
            " During the deposition, the DC bias is set at -440 V using the power of 56 W, and the substrate temperature is kept at 106 \u2070C. The film thickness of 140 nm is obtained after 40 min deposition time. Raman spectrometer (Jobin Yvon T64000) with Ar+ excitation source (wavelength = 514.5 nm) is used to study the DLC film structures and the data is analyzed by LabSpec software (version 3.03t). The DLC electrode is then inserted in a PDMS microfluidic chip [4]. The microchannel used as the electrochemical cell has a depth of 70 m and the overall volume of 2.24 \u03bcL. As shown in Fig. 1, the three-electrode chip consists of a DLC working electrode (1 mm2), a Ag wire reference electrode, and a sputtered Au on Si counter electrode (3 mm2). Cyclic voltammetry (CV) of 10 mM K3Fe(CN)6 in 1 M KNO3 is employed to study electrode responses in the cell. The solution is loaded to the microchanel using a syringe. The potential is swept from 0.7 V to -0.3 V and back to 0.7 V at scan rates ranging from 50 - 500 mV/s. The following reagents are used for covalent coupling of enzymes onto the electrode: glucose oxidase (GOx) from Aspergillus niger (E"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002979_20110828-6-it-1002.02676-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002979_20110828-6-it-1002.02676-Figure1-1.png",
        "caption": "Fig. 1. Quad-rotor aircraft representation",
        "texts": [
            "00 \u00a9 2011 IFAC 13582 10.3182/20110828-6-IT-1002.02676 In this section is presented the dynamical model of the quad-rotor aircraft by using the Euler-Lagrange approach. 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 (see Fig. 1). Based on Castillo et al. (2005), we have used the next dynamical model: mx\u0308= u(cos\u03c8 sin \u03b8 cos\u03c6+ sin\u03c8 sin\u03c6), my\u0308= u(sin\u03c8 sin \u03b8 cos\u03c6\u2212 cos\u03c8 sin\u03c6), mz\u0308 = u cos \u03b8 cos\u03c6\u2212mg, \u03c8\u0308 = \u03c4\u03c8 , \u03b8\u0308 = \u03c4\u03b8, \u03c6\u0308 = \u03c4\u03c6 where u is the total thrust, m represents the mass of the aircraft, g denotes the gravitational constant. \u03c4\u03c8 , \u03c4\u03b8 and \u03c4\u03c6 are the yaw, pitch, and roll torques respectively. To maintain \u03c8i = 0 \u2200t > 0, the yaw control used is \u03c4\u03c8i = \u2212kp\u03c8\u03c8i \u2212 kv\u03c8\u03c8\u0307i (2) where kp\u03c8 and kv\u03c8 denote the proportional and derivative constants for this PD yaw control"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.10-1.png",
        "caption": "Figure 4.10 Common rib design",
        "texts": [
            " If this ring exists, the position of the bearing is determined by the actual ring width, that is, the width of the seal going inside the housing for 2\u20133mm. In order to increase stiffness and strength and for better heat withdrawal, especially in cast housings, ribs are applied. By the distribution of ribs, it is possible to improve casting conditions, facilitate contraction and reduce the occurrence of residual stresses. For a more uniform cross-section and better casting conditions, the common rib design is carried out pursuant to Figure 4.10, where all required dimension ratios are given. The shape of the gear body mainly depends on the operating conditions, service life, load, material and dimensions, size of the series and the technology available. When it is necessary for a pinion to be made from very high-quality and expensive material, it is worked separately and joined with the shaft either by press-fitted joints (Figure 4.11a), by spline joints or by key joints (Figure 4.11b). The design where the pinion is made in one piece with the shaft is recommended (Figure 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002275_amm.371.250-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002275_amm.371.250-Figure2-1.png",
        "caption": "Fig. 2. Catalyst soft program simulation for: a) the half-mold classic design; b) thin walls mold-right",
        "texts": [
            " In these papers is presented the study made with FDM system and \u201cDimensions sst\u201d machine only. As it may be seen from the Fig. 1 and after the \u201cCATALYST\u201d FDM system software simulations, the authors made, step by step, various design transformations concerning the industrial parts \u201dHalf-mold Lever\u201d connected with the specific capabilities of the RP technologies. There have been several attempts on simulations to determine suitable part deposition orientation for different objectives like dimensional accuracy, build time, support structure and his effect, etc. (Table 1 and Fig. 2). pressure some simulation tests using SolidWorks Simulation software were done for different materials and for different RP technologies (Table 2, Fig. 3 and Fig. 4). for this sort of parts are CT 4=0.36mm, CT5=0.5mm, CT6=0.7mm. Prototyping machines (naturally and the technologies too) the authors has built a number of \u201cHalfmold Lever\u201d at University of Applied Science of Aachen, Germany laboratories area. The used RP technologies was as follows; FDM-Fused Deposition Modeling, \u201dDimension sst\u201d machine; 3Dprintig, \u201dZ printer 450\u201d machine and PolyJet, Objet 30 machine"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001697_v10264-012-0025-0-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001697_v10264-012-0025-0-Figure4-1.png",
        "caption": "Fig. 4 The projection of point P from the forehead to the ground plane tool",
        "texts": [
            " If the value of touch is pointed, we can state that the first touch of the \"S\" is particularly inappropriate because it is connected with the tip of the plate where there is a danger of fracture due to shock. The contact line between \"UV\" and \"UT\", e.g. edge contact outside \"delicate\" contact with the cutting edge is acceptable. Possible, but not particularly suited is the shot with \"SV\". It might be said that a decisive influence on the situation has the angle of entry into plane of the workpiece \u2013 \"\u03b5\", which is the position adjustment tool to the workpiece. Analytical relations for the determination of contact can be derived from Fig. 4. The figure suggested cutting plate (front plate) and the ground plane Pr. From the figure it is possible to derive the distance of any point \"P\" face plate from the ground plane. Projection of point \"P\" to the base plane is the point \"R\". The front surface of the plate is generally inclined at angles \u03b3p and \u03b3f (see Fig. 4). Axis of miller are oriented perpendicular to the work surface. In the analysis we consider that the start-up plane is parallel to the direction of tool feed rate, while parallel to the axis of the milling head. Determination of the distance of point P on the face of the tool from PR: From Figure 4 it is possible to express the distance \" h\u010d \" between the point P on the head of the tool and its projection into the plane of the base. h\u010d = hp + tg\u03b3p + f \u00b7 tg\u03b3f (1) Point T is given by: removing the height hp = h (height removal at point T) and an additional angle (profile of the main cutting edge) = 90\u00b0 \u03c8r \u2013 \u03c7r substituting into the equation hp = h f = \u2013h \u00b7 tg\u03c8r (2) getting hT = h \u00b7 [tg\u03b3p \u2013 tg\u03c8r \u00b7 tg\u03b3f] (3) Point U: U is given hp = h f = \u2013h \u00b7 tg \u03c8r + fz \u00b7 cos\u03b5 (4) ht = h \u00b7 tg\u03b3p + (\u2013h \u00b7 tg\u03c8r + fz \u00b7 cos \u03b5) \u00b7 tg\u03b3f = = h \u00b7 tg\u03b3p \u2013 h \u00b7 tg\u03c8r + fz \u00b7 cos\u03b5 \u00b7 tg\u03b3f (5) ht = h[tg\u03b3p \u2013 tg\u03c8r \u00b7 tg\u03b3f] + sz \u00b7 cos\u03b5 \u00b7 tg\u03b3f (6) Point V is given: hp = 0 f = fz \u00b7 cos\u03b5 hv = hp \u00b7 tg\u03b3p + f \u00b7 tg\u03b3f (7) hv = fz \u00b7 cos\u03b5 \u00b7 tg\u03b3f (8) In determination of the place of the first contact with the workpiece tool let us assume that start-up plane is parallel to the direction of feed rate and also the axis of the tool"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003528_amr.706-708.1209-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003528_amr.706-708.1209-Figure1-1.png",
        "caption": "Fig. 1 Radial error characteristics",
        "texts": [
            " Features of grinding involute gears on a worm wheel gear grinding machine Usually, there are serious tooth-period errors in the process of grinding Involute gears on a worm wheel gear grinding machine. And tooth-period errors consist of radial errors caused by centre\u2013to\u2013centre spacing of the grindstone and the machined gear( such as a radial runout of grindstone, and lack of dynamic balance), tangential errors caused by circumferential position errors between the grindstone and the work-piece( such as transmission errors, a axial inching, helix errors of grindstone, etc). Referencing Fig. 1(a) and Fig. 2(a), the axial section can be treated as a rack. The process of grinding a gear with grindstone is regarded as the engaging process of the machined gear and the rack[3]. (1) In the initial position, the rack (grindstone) is tangent to machined gear tooth at point Q and point 'Q . The two points are just right at constant chord, which are considered as the start points of tooth profile error computation. (2) When the machined gear turns, the contact points on left and right tooth surface take the rightabout movement from point Q and 'Q ",
            "208, University of Auckland, Auckland, New Zealand-06/07/15,01:40:48) the dedendum circle. And when the grindstone turns anticlockwise, contact points of the two tooth surface move opposite direction. (3) If the machined gear turning angle is\u03d5 , then the left tooth surface contact point spread angle from point 'Q is equal to the right tooth surface contact point spread angle from point Q, and the relative spread lengths are equal too. Characteristics of radial errors of a worm wheel gear grinding machine Referencing Fig. 1(a), when the grindstone has a radial runout, errors of left and right tooth surface are caused, and the two errors are equal and have the same direction. For a single-head worm grindstone, a formula is given as the following. ' rr tt \u2206=\u2206 , )sin( rrr zet \u03d5\u03d5 +=\u2206 , )cos( ' rrr zet \u03d5\u03d5 +=\u2206 (1) Thereinto, re is a radial eccentric, z is gear tooth number, \u03d5 is the gear turning angle, r\u03d5 is the initial angle. If the spread length is thought as the longitudinal coordinate of the gear profile error curve, then it is easy to see that the gear profile error caused by a radial eccentric has a period of each grindstone turn and the wavelength is the base pitch. For the profile error is related to the relative position of the grindstone and the machined gear, the four situations are given as \u00b00 , \u00b090 , \u00b0180 , \u00b0270 \u3002The angle is the initial position of the grindstone setting, seeing Fig 1(b), (c), (d), (e). For example, o0 shows as high spot tool setting, and \u00b0180 shows as low spot tool setting, and so on. Four error curves have the same representation, and the different pattern. For instance, while \u00b00 tool setting, the high spot of grindstone grinding point Q and 'Q of the machined gear, which is the wave trough of the gear profile error. With the machined gear turning, point d\u2019 responds to point e, point d responds to point e\u2019, on which there are equal errors. ed tt \u2206=\u2206 ' , 'ed tt \u2206=\u2206 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure10-1.png",
        "caption": "Figure 10. Comparison of Normalised stresses with penetration & without penetration",
        "texts": [
            " Displacement and stress results with sub-modeling is compared with full model linear static analysis and observed the same results. This has given confidence that submodeling is properly done. FE model should be modeled without any penetration. This is a pre-requisite to carry out contact analysis otherwise it will lead to unrealistic results. Figure 8 shows the super element at RSFB region and Figure 9 shows the Penetration plot at RSFB region. It is observed that bottom region of the suspension mounting bracket is penetrating inside the external flitch. Figure 10 shows the comparison of stresses with and without penetration at RSFB region. It is observed from Figure 10 that unrealistic stress occurs at the location where penetration happens. Table 1 shows the comparison of stresses and displacements with and without penetration. It is observed from Table 1 that there is a significant difference in stresses. There is not much difference in displacements with and without penetration. Hence, FE model needs to prepared in such a way that no penetration should happen between the components Surface to surface contact is defined for the following contact pairs \u2022 RSFB bracket & External flitch \u2022 External Flitch & FSM \u2022 FSM & Internal flitch \u2022 Internal Flitch & RSFB Cross member Local non-linear analysis with contact has been carried out for the identified critical locations and compared the results between linear analysis and contact analysis"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003960_amr.652-654.1842-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003960_amr.652-654.1842-Figure1-1.png",
        "caption": "Fig. 1 Experimental Gear and grid bionic surface micro-morphology",
        "texts": [
            " In this work, the fatigue resistance characteristic of bionic surface distribution was attempted on gear surface, gridding form non-smooth units were designed and manufactured on gear surface through laser engraving technique and the fatigue resistance performance was discussed finally. Materials. Anti-fatigue performance was tested by CL-100 testing machine. The material of gears was 20Cr. One of the comparison samples was a pair of mesh gears with bionic surface and the other was a general pair. The pinion had 16 teeth and 24 for the bigger one (Fig. 1a). Gear teeth with bionic surface showed as Fig. 1b which was processed by CT-200 II numerical control laser carving machine with stationary laser lens. The micro-configuration was gridding forms in the experiment. The gears were fixed on a rotational platform and left-right shifted under the pulsed laser lens. The gridding configurations were carved on pitch circle position finally (showed in 1b). In order to obtain different gridding shapes, some system parameters must be set up first, such as pulse width, X-axis speed, current intensity, and so on. Table 1 showed 9 types gridding shapes and their processing parameters. Among them, the microscope image of only sample No.1 was illustrated as Fig.1b. Fatigue Tests. Fatigue tests were conducted by CL-100 gear fatigue testing machine with power distributional closed loop. Every pair of gear samples was kept in no-load running condition for 2 hours. Then, change the lubricating oil and running in load. This testing course divided into two 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: 141.117.125.76, Ryerson University Lib, Toronto-26/04/15,22:57:30) phases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.13-1.png",
        "caption": "Figure 4.13 Typical design of a cast gear: (a) with crossed spokes, (b) with I-shaped spokes",
        "texts": [
            " In the individual production of gears with diameter d 400mm, welding is common. The choice of procedure for smalland medium-scale production depends on the results of the optimization. However, the problems of making the casting tools (moulds) and blacksmithing tools (dies) grow with the size of the gear, hence a growing tendency towards welding appears. All of this relates to the working of the gear body. Teeth are cut, heat-treated and finally subjected to a finishing process, except in sand casting where they are usually not subjected to any finishing working. Figure 4.13 shows typical designs of a gear manufactured by casting in which the rim is connected to the hub with crossed spokes if d 1000mm and width b< 200mm (Figure 4.13a) and with I-shaped spokes if d> 1000mm and width b> 200mm (Figure 4.13b). Forged wheels are made solid or cored with round holes (Figure 4.14). The cored type is lighter, but requires more machining. The type without holes is simpler to work but is very heavy if of large width, and it does not allow homogenous mechanical properties to be obtained in the teeth after heat treatment. For a more convenient clamping of wheels on the machine tool, the web of the disc should be drilled between the rim and the hub. Sometimes large diameter holes are drilled to reduce the weight of the wheels"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure3-1.png",
        "caption": "Fig. 3 Medical robotic arm that is attached to an operating table",
        "texts": [
            " When the doctor push the male side, the spring will be press and the O-ring will be the lock part for this connection. The key point of this kind of design is to make sure that the O-ring has the right size for locking. Other structures are just the same as pneumatic connector. C. Connection based on the Leverage Theory The typical example of this kind of connector, \u201cMedical robotic arm that is attached to an operating table\u201d [6], is a part of the famous minimally invasive surgery system device, da Vinci Si HD Surgical System [7]. The structure is demonstrated as the following Fig.3. As is shown in the Fig.3, the quick connector 2 has a slot 3 that receives a pin 4 of the front loading tool driver 1. The pin 4 can be released by depressing a spring biased lever 5. When the surgeon put a handle in the position 2, and the connector 3 will be pushed into 1, and groove goes to the pin 4, and then based on the Leverage Theory, the spring on 5 will be pressed, the pin will move into the groove on 3, so the handle is connected to the female machine. The handle will be released by a simple press on 5. While designing this kind of structure, the press forth and the length of the level would be highly considered to make sure the reliability of the connection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003445_ht2012-58139-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003445_ht2012-58139-Figure11-1.png",
        "caption": "Fig. 11: Equivalent von-Mises stress distribution on the welded plate.",
        "texts": [
            " This can be explained on the basis of end-effect. The plots for directional deformation are not shown here, but it is seen that the symmetry face ABCD shows no deformation in the x-direction. This confirms correct set up of the model in the FE solver, as the symmetry boundary condition is upheld. Such contour plots help us visualize the nature of the deformation of the welded joint and thus can prove significant from a design perspective. We note that the total deformation of the structure is represented by 2 2 2 x y z\u03d1 \u03d1 \u03d1+ + . Figure 11 highlights the equivalent von-Mises stress distribution on the welded plate, calculated from the principal stresses as 2 2 2 1 2 2 3 3 10.5*{( ) ( ) ( ) }\u03c3 \u03c3 \u03c3 \u03c3 \u03c3 \u03c3 \u03c3\u2032 = \u2212 + \u2212 + \u2212 (25) where, 1\u03c3 , 2\u03c3 and 3\u03c3 are the principal stress components. The von-Mises equivalent stress is used to predict yielding of the welding joint, i.e. when the von-Mises stress becomes greater than the yield stress the structure undergoes yielding due to plasticity. Figure 12 shows the equivalent von-Mises stress on the butt joint weld plane ABCD"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.10-1.png",
        "caption": "Fig. 3.10 Four bar linkage",
        "texts": [
            " For the dual string pendulum the radius of gyration, , of the rod and point masses about the vertical axis through its centre is c, so Eq. 3.11 reduces to Eq. 2.13. Hence, for small amplitudes the time of swing is independent of the length of the rigid massless rod. In mechanical engineering terms the trapezium pendulum is a Watt\u2019s linkage, developed in 1784 (Tre\u0301baol 2008), and sometimes called Watt\u2019s parallel motion (Dunkerley 1910). Watt\u2019s linkage is an example of a four bar linkage. In a four bar linkage the four links are pivoted together so that they move within a plane, as shown schematically in Fig. 3.10, These links are: a static bar, regarded as being fixed in space, a pilot crank, a connecting rod, and a rocker. A coupler may be attached to one of the moving links. A four bar linkage has one degree of freedom since one parameter, for example the angle between the static bar and the pilot crank, is needed to specify its configuration. The trapezium pendulum as a form of Watt\u2019s linkage is shown in Fig. 3.11. The pilot crank and rocker of a four bar linkage are the two rigid massless rods, length l1, shown in the figure"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002459_urai.2012.6463047-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002459_urai.2012.6463047-Figure1-1.png",
        "caption": "Fig. 1. Finishing robot system",
        "texts": [
            " In order to realize the system, we are developing programming less force controlled finishing robot system which is applied environment and task motion framework technology [1]. Information to carry out finishing works by the robot is put separately in tools, works, and objects in the working environment under environment and task motion framework technology by using RFID tags. The finishing robot system automatically generates robot motion programs by using information obtained from the environment. We aim to develop a robot system which can execute appropriate finishing works (Fig.1). In order that the robot handles the workpiece and carries out a finishing work automatically, position and posture information of a workpiece is needed. Camera and laser range finder are often used to measure the position and posture of an object [2]. Almost all workpieces in the finishing work environment are metal object. Because of metallic luster, accurate position and posture measurement is difficult by using camera and laser range finder. Furlani proposed object position and posture estimation method by using RFID and/or barcode tags on the assumption that these tags are read by human [3]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001959_9781119969242.ch7-Figure7.27-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001959_9781119969242.ch7-Figure7.27-1.png",
        "caption": "Figure 7.27 Inverting Schmitt trigger and wave-shaping circuit",
        "texts": [],
        "surrounding_texts": [
            "The inverters control logic can be implemented using an analog circuit and its complete block diagram is shown in Figure 7.24. The other method is by using advanced microprocessors, microcontrollers, and digital signal processor to implement the control schemes. In Figure 7.24, the supply is taken from a single-phase power supply and converted to 9-0-9Vusing a small transformer. This is fed to the phase shifting circuit, shown in Figure 7.25, to provide an appropriate phase shift for operation at various conduction angles. The phase shifted signal is then fed to the inverting/non-inverting Schmitt trigger circuit and waveshaping circuit (Figures 7.26 and 7.27). The processed signal is then fed to the isolation and driver circuit shown in Figure 7.28, which is then finally given to the gate of IGBTs. There are two separate circuits for upper and lower legs of the inverter. The power circuit can be made up of IGBT, with a snubber circuit consisting of a series combination of a resistance and a capacitor with a diode in parallel with the resistance."
        ]
    },
    {
        "image_filename": "designv11_100_0002760_imece2013-63365-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002760_imece2013-63365-Figure1-1.png",
        "caption": "Figure 1. Geometric conditions of non-conjugate surface contact of the gear pair",
        "texts": [],
        "surrounding_texts": [
            "NON-CONJUGATE TOOTH SURFACES 2.1 Theoretical background for the TE calculation The non-conjugate tooth surfaces in this paper are the surfaces of which the gearing law is not satisfied. If the nonconjugate tooth surfaces are combined together, the normal vector and the difference of two velocity vectors of pinion and gear is not always orthogonal. However for any arbitrary tooth surfaces, the directions of the normal vectors are always equal at the contact point. \ud835\udc41\ud835\udc56 \u00d7 \ud835\udc41\ud835\udc5c = 0 (1) The location of the contact point must be placed properly at the installed state. Being expressed in mathematical formulas, these are as follows: \ud835\udc56\ud835\udc5e = \ud835\udc56\ud835\udc5c + \ud835\udc5c\ud835\udc5e (2) Where \ud835\udc56\ud835\udc5e is expressed with the functions of curved surface parameters (\ud835\udc4e, \ud835\udc4f) and a mechanical parameter (\ud835\udf031) of the pinion surface as \ud835\udc56\ud835\udc5e (\ud835\udc4e, b, \ud835\udf031); \ud835\udc5c\ud835\udc5e is expressed with the functions of curved surface parameters (\ud835\udc50, \ud835\udc51) and a mechanical parameter (\ud835\udf032) of the gear surface as \ud835\udc5c\ud835\udc5e (\ud835\udc50, \ud835\udc51, \ud835\udf032); and \ud835\udc56\ud835\udc5c as a position vector between reference points of the pinion and gear is a prefixed constant value when the gears are installed. In the worm gear system, \ud835\udc56, \ud835\udc5c are generally the center points, respectively, located on the shortest paths between two different axes and \ud835\udc56\ud835\udc5c is set to be the center line of the shortest distance between two axes. When the surface of each gear is arithmetically defined, the normal vector can be obtained from the following formulas: \ud835\udc41\ud835\udc56 = \u2202\ud835\udc56\ud835\udc5e \u2202\ud835\udc4e \u00d7 \u2202\ud835\udc56\ud835\udc5e \u2202b (3) \ud835\udc41\ud835\udc5c = \u2202\ud835\udc5c\ud835\udc5e \u2202\ud835\udc50 \u00d7 \u2202\ud835\udc5c\ud835\udc5e \u2202d (4) Accordingly, \ud835\udc41\ud835\udc56 is expressed with the functions of parameters \ud835\udc4e, \ud835\udc4f and \ud835\udf031 . In short, \ud835\udc41\ud835\udc56 (\ud835\udc4e, \ud835\udc4f, \ud835\udf031) . \ud835\udc41\ud835\udc5c is expressed with the functions of parameters \ud835\udc50, \ud835\udc51 and \ud835\udf032 . In brief, \ud835\udc41\ud835\udc5c (\ud835\udc50, \ud835\udc51, \ud835\udf032) . From formula (1), two independent algebraic equations can be obtained as shown below \ud835\udc391(a, b, \ud835\udc50, \ud835\udc51, \ud835\udf031 , \ud835\udf032) = ( \ud835\udc41\ud835\udc56 \u00d7 \ud835\udc41\ud835\udc5c ) \u2022 \ud835\udc56 = 0 (5) \ud835\udc392(a, b, \ud835\udc50, \ud835\udc51, \ud835\udf031 , \ud835\udf032) = ( \ud835\udc41\ud835\udc56 \u00d7 \ud835\udc41\ud835\udc5c ) \u2022 \ud835\udc57 = 0 (6) From formula (2), three independent algebraic equations can be obtained as shown below. \ud835\udc393(a, b, \ud835\udc50, \ud835\udc51, \ud835\udf031 , \ud835\udf032) = (\ud835\udc56\ud835\udc5e \u2212 \ud835\udc56\ud835\udc5c \u2212 \ud835\udc5c\ud835\udc5e ) \u2022 \ud835\udc56 = 0 (7) \ud835\udc394(a, b, \ud835\udc50, \ud835\udc51, \ud835\udf031 , \ud835\udf032) = (\ud835\udc56\ud835\udc5e \u2212 \ud835\udc56\ud835\udc5c \u2212 \ud835\udc5c\ud835\udc5e ) \u2022 \ud835\udc57 = 0 (8) \ud835\udc395(a, b, \ud835\udc50, \ud835\udc51, \ud835\udf031 , \ud835\udf032) = (\ud835\udc56\ud835\udc5e \u2212 \ud835\udc56\ud835\udc5c \u2212 \ud835\udc5c\ud835\udc5e ) \u2022 \ud835\udc58 = 0 (9) From formulas (5) to (9), when \ud835\udf031 is set as a free parameter and curved surface parameters \ud835\udc4e, \ud835\udc4f, \ud835\udc50, \ud835\udc51 are eliminated, \ud835\udf032 becomes the function of \ud835\udf031. By using these, TE is calculated by using the following formula: \ud835\udc47\ud835\udc38(\ud835\udf031) = \ud835\udf032(\ud835\udf031) \u2212 \ud835\udf0320 (10) where, \ud835\udf0320 = \ud835\udc54\ud835\udf031. 2.2 The relationship between TE and TI When the tooth surface moves in the meshing region, a pair of gears starts to contact (contact-1) and when the surface gets out of the region, meshing is finished. When contact-1 moves as much as the pitch angle (\ud835\udf031 \ud835\udc5d = 2\ud835\udf0b \ud835\udc4d1 ), a pair of rear gears starts to contact (contact-2). When TE occurs, the phenomenon of teeth interference or separation occurs at contact-2. The detailed explanation is as shown below. Fig. 2 illustrates the case when contact-1 entered the meshing region and moves further as much as the pitch angle. TI occurs at the time that contact-2 just entered the meshing region. When the front pinion(driving) surface rotates as much as \ud835\udf031 from the reference line of the rotation angle, the front gear(driven) surface rotates as much as \ud835\udf032(\ud835\udf031). Since the rear pinion surface is placed behind as much as \ud835\udf031 \ud835\udc5d , its rotation angle is \ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d . Accordingly, the rear gear surface moves as much as \ud835\udf032(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d) from the reference line of the rotation angle. Let\u2019s take the rear gear surface at \ud835\udf032(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d) as the theoretical rear gear surface. Then the rotation angle between the front gear surface and the theoretical rear gear surface \ud835\udf032 \ud835\udc5d is established as follows: Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 09/01/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use 3 Copyright \u00a9 2013 by ASME \ud835\udf032 \ud835\udc5d = \ud835\udf032(\ud835\udf031) \u2212 \ud835\udf032(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d ) (11) But the actual rear gear surface is placed behind as much as the gear pitch angle ( \ud835\udf0320 \ud835\udc5d = 2\ud835\udf0b \ud835\udc4d2 ). When formula (10) is substitute into formula (11), the difference of the rotation angle of the theoretical rear gear and the actual rear gear is established as follows: \ud835\udf032 \ud835\udc5d \u2212 \ud835\udf0320 \ud835\udc5d = \ud835\udc47\ud835\udc38(\ud835\udf031) \u2212 \ud835\udc47\ud835\udc38(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d) (12) where, \ud835\udf0320 \ud835\udc5d = \ud835\udc54\ud835\udf031 \ud835\udc5d . If \ud835\udf032 \ud835\udc5d = \ud835\udf0320 \ud835\udc5d (or \ud835\udc47\ud835\udc38(\ud835\udf031) = \ud835\udc47\ud835\udc38(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d ) ), while the surfaces of the front pinion and front gear are being meshed, those of the rear pinion and rear gear start to be meshed smoothly and double contacts are developed without any troubles. For the reason that gears are actually placed at equal intervals as much as the pitch angle \ud835\udf0320 \ud835\udc5d , if \ud835\udf032 \ud835\udc5d < \ud835\udf0320 \ud835\udc5d (or \ud835\udc47\ud835\udc38(\ud835\udf031) < \ud835\udc47\ud835\udc38(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d)), the rear pinion surface placed ahead of the actual rear gear. Accordingly, the interference between the rear pinion surface and the actual rear gear surface occurs. In details, while the gear is inserted to the radial direction, the edge of the gear hits the pinion surface to the radial direction and moves in. Because this directly causes a rattle noise, and even damage of the tooth surface, this must be eliminated upon design. Fig. 3 represents TE occurring in the contact region when the pinion rotates. TE is repeated at a cycle which is the pitch angle because the front tooth surface and the rear tooth surface are placed at the interval as much as the pitch angle. Since the rotation angle in the contact region is greater than the pitch angle, double contacts occur at some region. In the \ud835\udc4f~\ud835\udc50 section in Fig. 3 as shown above, meshing is led by contact-1 and the gears are meshed and move by the contact at a single point. When contact-1 is at point \ud835\udc50 (\ud835\udf031c), contact-2 moves in the meshing region. At the time, TEs at contact-1 and contact-2 are \ud835\udc50\ud835\udc501 = \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc50) , and \ud835\udc50\ud835\udc4e2 = \ud835\udc47\ud835\udc38(\ud835\udf031c \u2212 \ud835\udf031 \ud835\udc5d), respectively. In Fig. 3, because \ud835\udc50\ud835\udc501 < \ud835\udc50\ud835\udc4e2 or \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc50) < \ud835\udc47\ud835\udc38(\ud835\udf031c \u2212 \ud835\udf031 \ud835\udc5d), a tooth impact occurs at contact-2 (Refer to Fig. 2). After the impact, as the rotations of gears increase discontinuously as much as \ud835\udc501\ud835\udc4e2, teeth are separated at contact-1 and the gears move to \ud835\udc51 while meshing is led by contact-2. When they go past \ud835\udc51, contact-1 gets meshing again and it is developed until \ud835\udc52 , where the meshing region ends. At the time, contact-2 is the state of the tooth surfaces being separated. At point \ud835\udc52, (\ud835\udf031\ud835\udc52) the TEs at contact-1 and contact2 are \ud835\udc52\ud835\udc521 = \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc52), and \ud835\udc52\ud835\udc4f2 = \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc52 \u2212 \ud835\udf031 \ud835\udc5d), respectively, when the gap between tooth surfaces of contact-2 is largest. The maximum value of the gap (\ud835\udc3amax ) is calculated in the following formula: \ud835\udc3amax = \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc52) \u2212 \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc52 \u2212 \ud835\udf031 \ud835\udc5d ) (13) After that, while contact-1 ends out of the meshing region, the gap (\ud835\udc3amax ) at point \ud835\udc52 is filled, the front impact occurs between the tooth surfaces. Consequently, Fig. 4 represents TE actually occurring as shown in Fig. 3 with a thick line. To put it shortly, if \ud835\udc47\ud835\udc38(\ud835\udf031) < \ud835\udc47\ud835\udc38(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d), TI occurs at point \ud835\udc50 leading the impact between the edge of the gear and the pinion surface to the radial direction and the teeth Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 09/01/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use 4 Copyright \u00a9 2013 by ASME separation occurs at point \ud835\udc52 leading the frontal impact between the tooth surfaces. TI at point \ud835\udc50 causes fatal damage to the tooth surface and creates rattle noise due to the high impact speed to the radial direction. Therefore, TI must be completely removed. Since the frontal impact occurs on the tooth surface due to the gap at point \ud835\udc52, the impact speed is not higher than that at point \ud835\udc50. TE works as a vibration source leading not only the rattle noise by tooth impact but also the whine noise of the gear system with the form of displacement excitation at the contact region. To minimize the whine noise, the design strategy, therefore, must be made to the direction of minimizing the size of the frequency spectrum of the periodic function of the TE calculated from Fig. 4. Fig. 5 shows the case that the teeth separation occurs without TI. In formula (12), when \ud835\udc47\ud835\udc38(\ud835\udf031) > \ud835\udc47\ud835\udc38(\ud835\udf031 \u2212 \ud835\udf031 \ud835\udc5d), \ud835\udf032 \ud835\udc5d > \ud835\udf0320 \ud835\udc5d . As the actual rear gear surface is placed ahead of the rear pinion surface, no contact occurs at contact-2, but the gap between the tooth surfaces occurs. While passing through the meshing region, contact-1 continuously leads the contact and contact-2 is developed while the tooth surfaces are separated. At the moment that the tooth contact of the front pinion and the front gear at contact-1 are ended, frontal impact of the tooth surfaces at contact-2 occurs. Fig. 6 shows the distribution of TEs. Contact-1 leads meshing at point \ud835\udc4f to point \ud835\udc52 via \ud835\udc50 without any sudden change. The region of \ud835\udc4f~\ud835\udc50 is developed at the contact of a single point. At point \ud835\udc50, contact-2 enters the meshing region. The TEs of contact-1 and contact-2 are \ud835\udc50\ud835\udc501 = \ud835\udc47\ud835\udc38(\ud835\udf031\ud835\udc50) and \ud835\udc50\ud835\udc4e2 = \ud835\udc47\ud835\udc38(\ud835\udf031c \u2212 \ud835\udf031 \ud835\udc5d) , respectively. In Fig. 6, \ud835\udc50\ud835\udc501 > \ud835\udc50\ud835\udc4e2 or \ud835\udc47\ud835\udc38(\ud835\udf031c) > \ud835\udc47\ud835\udc38(\ud835\udf031c \u2212 \ud835\udf031 \ud835\udc5d ). The gap between tooth surfaces of contact-2 is formed as much as \ud835\udc4e2\ud835\udc501 (Refer to Fig. 5). Accordingly, moves into the meshing region while the teeth are separated. Meshing is still led by contact-1, which moves until \ud835\udc52 where the meshing region ends. At point \ud835\udc52 (\ud835\udf031\ud835\udc52), the gap between the tooth surface of contact-1 and that of contact-2 may be obtained from formula (13). Fig. 7 represents TEs actually occurring in Fig. 6 with a thick line. According to the meshing mechanism as shown above, it can be found that the gear system made in combination with the non-conjugate tooth surface causes TI and TEs. To commercialize the gear system in combination with the nonconjugate tooth surface, TI must be removed and TEs must be minimized. To remove TI, it, therefore, must be satisfied that \ud835\udc50\ud835\udc501 > \ud835\udc50\ud835\udc4e2 at point \ud835\udc50, or \ud835\udc47\ud835\udc38(\ud835\udf031c) > \ud835\udc47\ud835\udc38(\ud835\udf031c \u2212 \ud835\udf031 \ud835\udc5d). When the gap (\ud835\udc3amax ) between the tooth surfaces is minimized at point \ud835\udc52, the TE is minimized."
        ]
    },
    {
        "image_filename": "designv11_100_0002035_978-3-642-28956-9_8-Figure8.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002035_978-3-642-28956-9_8-Figure8.3-1.png",
        "caption": "Fig. 8.3 Membership functions of the input/output variables",
        "texts": [
            " The purpose of the FLC is to find a control input s such that the current velocity vector v is able to reach the velocity vector vc this is denoted as (8.8): lim t!1 vc vk k \u00bc 0 \u00f08:8\u00de The inputs variables of the FLC correspond to the velocity errors obtained of (8.6) (denoted as ev and ew: linear and angular velocity errors respectively), and 2 outputs variables, the driving and rotational input torques s (denoted by F and N respectively). The initial MF are defined by 1 triangular and 2 trapezoidal functions for each variable involved. In future work the shape of the MF will be selected by the algorithm as part of the optimization. Figure 8.3 depicts the MFs in which N, C, P represent the fuzzy sets (Negative, Zero and Positive respectively) associated to each input and output variable. Table 8.1 shows the upper and lower limits of the used MF. The rule set of the FLC contain nine rules, which govern the input\u2013output relationship of the FLC and this adopts the Mamdani-style inference engine. We use the center of gravity method to realize defuzzification procedure. In Table 8.2, we present the rule set whose format is established as follows: Rule i : If ev is G1 and ew is G2 then F is G3 and N is G4 \u00f08:9\u00de Where G1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003545_20130619-3-ru-3018.00528-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003545_20130619-3-ru-3018.00528-Figure1-1.png",
        "caption": "Fig. 1. a) direction-changing, b) direction-preserving saturation (left: control vector before saturation, right: after saturation)",
        "texts": [
            " For 0 < \u03c1k \u2212,t,i \u2264 \u03c1k +,t,i < 1, where is a partial failure in the actuator and information loss carried by sat(vt,i). For %\u2212,i = %+,i there are no constraints imposed on the control system and uF t,i = vt,i, the case %\u2212,i > 0 corresponds to partial failure, and %\u2212,i = 0 corresponds to outage case. Actuator failure may give rise to changes in direction between computed v t and applied u t control vectors. 978-3-902823-35-9/2013 \u00a9 IFAC 1810 10.3182/20130619-3-RU-3018.00528 Depending on the method of imposing constraints on the control vector one can observe directional change, illustrated in Fig. 1a in the case of cut-off saturation (for future reference), which is not present when saturation is performed according to imposed constraints (dashed lines) with constant direction, Fig. 1b (for future reference: DP). Let two-input two-output system be not coupled and both loops be driven by separate I controllers (with no cross-coupling in the linear case). The output vector y t is to track reference vector of two sinusoid waves, what corresponds to drawing a circular shape in the (y1, y2) plane [1]. As it can be seen in the Fig. 2a, the system with no constraints performs best, whereas in the case of cutoff saturation of both the elements of control vector (Fig. 2b) the tracking performance is poor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.72-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.72-1.png",
        "caption": "Figure 2.72. Definition of the coordinate",
        "texts": [
            "210) where lss\u00bd is the matrix of self- and mutual inductances of the stator windings, lsr\u00bd is the matrix of mutual inductances between the stator and rotor windings, and lrr\u00bd is the matrix of self- and mutual inductances of the rotor windings. Considering the symmetrical structure of the rotor, it results that the self- and mutual inductances of the stator and rotor windings, respectively, are constant, while the mutual inductances between the two windings depend on the rotor position given by the angle u (Figure 2.72): u \u00bc vrt \u00fe u0 (2.211) where u0 is the value of u at the instant t \u00bc 0. Substituting vr from (2.206) in (2.211) achieve the relationship that describes the dependency of u in terms of the slip s and the angular velocity of the rotor field vs, in el. rad/s: u \u00bc 1 s\u00f0 \u00devst \u00fe u0 (2.212) Therefore, laa \u00bc lbb \u00bc lcc \u00bc Laa and lab \u00bc lbc \u00bc lca \u00bc Lab for the stator windings, and lAA \u00bc lBB \u00bc lCC \u00bc LAA and lAB \u00bc lBC \u00bc lCA \u00bc LAB for the rotor windings, respectively. Hence, lss\u00bd \u00bc Laa Lab Lab Lab Laa Lab Lab Lab Laa 2 664 3 775 (2",
            "213b) Assuming a sinusoidal distribution of the magnetic flux in the air gap, the mutual inductances between the stator and rotor windings have a sinusoidal variation in terms of u, which defines the rotor position. Therefore, the mutual inductance between a stator winding and a rotor winding is maximum when the axes of the two windings coincide, decreases as the angle u increases and becomes zero when the axes are perpendicular, decreases again as the angle increases and reaches a maximum negative value when the axes are in opposition and so forth. Based on these aspects and considering the angles defined in Figure 2.72, it results lsr\u00bd \u00bc LaAcos u LaAcos u \u00fe 2p 3 LaAcos u 2p 3 LaAcos u 2p 3 LaAcos u LaAcos u \u00fe 2p 3 LaAcos u \u00fe 2p 3 LaAcos u 2p 3 LaAcos u 2 6666666664 3 7777777775 (2.213c) lrs\u00bd \u00bc lsr\u00bd t \u00bc LaAcos u LaAcos u 2p 3 LaAcos u \u00fe 2p 3 LaAcos u \u00fe 2p 3 LaAcos u LaAcos u 2p 3 LaAcos u 2p 3 LaAcos u \u00fe 2p 3 LaAcos u 2 6666666664 3 7777777775 (2.213d) where LaA is the maximum value of the mutual inductances. From equations (2.210), (2.213a), (2.213b), (2.213c), and (2.213d), it results the following expressions for the total flux linkage in the stator winding a and in the rotor winding A: ca \u00bc Laaia \u00fe Lab ib \u00fe ic\u00f0 \u00de \u00fe LaA iA cos u \u00fe iB cos u \u00fe 2p 3 \u00fe iC cos u 2p 3 cA \u00bc LAAiA \u00fe LAB iB \u00fe iC\u00f0 \u00de \u00fe LaA ia cos u \u00fe ib cos u 2p 3 \u00fe ic cos u \u00fe 2p 3 (2",
            "216) can be obtained also for the total flux linkage in the windings of the b and c phases and of the B and C phases, respectively. 2.2.2.2 The d\u2013q Transformation. As in the case of the synchronous machine, the equations of the induction motor can be simplified by applying the change of variables. If the d\u2013q frame attached to the rotor was chosen for the synchronous generator, for the induction motor it is more convenient to use a d\u2013q frame rotating with the synchronous speed. The q-axis is assumed to be p=2 ahead of the d-axis in the direction of rotation [1] (Figure 2.72). Therefore, two Park transformation matrices are required: one to transform the quantities and equations attached to the stator windings and the other to transform the quantities and equations attached to the rotor windings. Let us be the angle between the d-axis and the fixed stator phase a-axis at the instant t, and ur be the angle between the d-axis and the rotational rotor phase A-axis at the same instant. Selecting as time reference the instant at which the three axes (d, a, and A) overlap then u0 \u00bc 0 and, according to equation (2.211) and Figure 2.72, gives u \u00bc vrt ; us \u00bc vst ur \u00bc us u \u00bc vs vr\u00f0 \u00det \u00bc svst ( (2.217) Therefore, the transformation matrices for the stator (a, b, c) and rotor (A, B, C) phase quantities in (d, q, 0) quantities, known as the Park transformation matrices for the induction motor, are Ps\u00bd \u00bc 2 3 cos us cos us 2p 3 cos us \u00fe 2p 3 sin us sin us 2p 3 sin us \u00fe 2p 3 1 2 1 2 1 2 2 666666664 3 777777775 (2.218) and Pr\u00bd \u00bc 2 3 cos ur cos ur 2p 3 cos ur \u00fe 2p 3 sin ur sin ur 2p 3 sin ur \u00fe 2p 3 1 2 1 2 1 2 2 666666664 3 777777775 (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003768_ht2012-58524-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003768_ht2012-58524-Figure2-1.png",
        "caption": "Figure 2 Schematic of test set-up using embedded thermocouples and a stack of washers, after [16].",
        "texts": [
            " This is then used to predict the thermal contact behaviour between flat washers and fly cut washers. Validation of the FE model is carried out by comparing its predicted TCC values with experimental measurements. The experimental and modelling approaches adopted in this work is an extension of those developed by Woodland [16]. All the experiments in this study examine PE16, a high temperature alloy used in turbine casings. The specimens were tested by a split tube instrumented technique developed by Woodland et al [17]. It is a steady-state technique and the set up is shown in Figure 2. Thermocouples embedded in the steel cylinders measure the temperature profile along the rig\u2019s length. The thermocouples were calibrated to within 0.1 o C relative to each other. A band heater heats the top steel cylinder and the bottom cylinder is water-cooled. Band heater for this study was controlled at 100\u00b0C. Multiple specimens were stacked in between the steel cylinders to provide the necessary temperature drop to calculate the thermal contact conductance. The use of a multiple stack reduces the errors arising from the uncertainty of thermocouple readings; see Woodland et al [17] for a full explanation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure12-1.png",
        "caption": "Figure 12. Input Force to the Articulation Sternoclavicularis",
        "texts": [
            " As the human muscle cannot exercise infinite amount of force, it can be said that the maximum amount of muscle force of the trainee limits the range of scapula\u2019s backward motion. 5 F and 6 F from the clavicle and the rib are almost constant in both Fig.9 and 11, because 3 \u03b8 is not changed, and the input force F is parallel to the clavicle and the rib. IV. SHOULDER FORCE WHEN EXTERNAL INPUT IS GIVEN The force of shoulder muscle is analyzed when three different outside forces are given. A. Input to Articulatio Acromioclavicularis (Analysis 1) External input is added at the articulation sternoclavicularis (Fig. 12 and Table V). a F represents arm\u2019s weight. b F represents the external input. The force of shoulder muscle, and the force of clavicle and the rib are shown in Fig. 13. B. Input to Scapulothoracic Joint (Analysis 2) External input c F is added at the scapulothoracic joint (Fig. 14 and Table VI). The force of shoulder muscle, the force of clavicle and the rib are shown in Fig. 15. C. Input to Articulatio Acromioclavicularis and Scapulothoracic Joint (Analysis 3) External input is added evenly at the articulation sternoclavicularis ( b F ) and the scapulothoracic joint ( c F ) (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.44-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.44-1.png",
        "caption": "Fig. 2.44 4PaPPaR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\P\\\\Pa\\\\R",
        "texts": [
            " 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig. 2.49) R\\Pa||Pa\\kR (Fig. 2.22j) The first and the last revolute joints of the four limbs are parallel 38"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003850_amm.220-223.1040-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003850_amm.220-223.1040-Figure1-1.png",
        "caption": "Fig. 1 Diagram of the space vector of high-speed PMSM",
        "texts": [
            " The control rule of fuzzy controller can be adjusted according to the input converting the expert knowledge to automatic control strategy. Especially when the information processing changes suddenly, fuzzy control plays a vital role to improve the dynamic performance. MATLAB simulation is done for sensorless grey prediction fuzzy DTC system of the high-speed PMSM in this paper. The simulation results show that the sensorless grey prediction fuzzy DTC has higher control accuracy and faster response than conventional DTC. Mathematical model of the high-speed PMSM The space vector of high-speed PMSM is shown in Fig. 1. In the two Cartesian coordinate system in Fig. 1: \u03b1, \u03b2 are the stator stationary coordinate system. \u03b1 axis is coincident with A axis of stator winding. d, q are the rotation coordinate system for rotor. d axis is coincident with the direction of rotor flux linkage and rotates at synchronous speed \u03c9r in counter-clockwise. The angle between the two coordinate systems is \u03b8. 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_100_0001842_9781118561799.ch1-Figure1.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001842_9781118561799.ch1-Figure1.8-1.png",
        "caption": "Figure 1.8. The principal of molecular imprinted polymer fabrication",
        "texts": [
            " Methods for cell immobilization by trapping them inside a polymeric network or by grafting using functionalized support with antibodies by antigen-antibody interactions are the most common because they are mild and enable the micro-organisms\u2019 metabolic activity to be preserved. 1.3.7. A biomimetic approach: molecularly imprinted polymers (MIPs) Molcular imprinting can be defined as the formation of specific recognition sites (with bonding or catalytic properties) in a material from its interaction with a \u201ctemplate\u201d molecule, where it directs the positioning and orientation of the material\u2019s structural components by an auto-assembly mechanism (see Figure 1.8). The material itself can be composed of a linear sequence or a polymerical sequence (organic or inorganic silicon gel-based imprinted MIPs) or a 2D surface network (grafted monolayers). MIPs have a number of advantages when compared with natural biomolecules: they can be prepared for specific molecules (the best results have been obtained for molecules with a molecular weight of between 200 and 1,200 Da), they are stable for a large range of pH, pressure and temperature, can function as organic solvents, are insoluble in water, can easily create multisensor networks and are compatible with micro-production techniques"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.14-1.png",
        "caption": "Fig. 7.14 The first and second of four I-K solutions for a Stanford arm",
        "texts": [],
        "surrounding_texts": [
            "To further reduce the dimension of a combined isometric embedding given in (7.55), let us now pay more attention to the last link, i.e., the kinematically \u201cbusiest\u201d link of an n-joint open serial-chain robot. Obviously, the isometric embedding of link n, zn = \u03b6n(q) can embed the C-manifold Mn of link n into R 9 by (7.54) with i = n. Intuitively, because any joint movement will contribute a non-trivial part to the last link motion, the dimensions of Mn for the last link and Mc for the entire robot should both be equal to n. Now the question is can the mapping zn = \u03b6n(q) of the last link also embed the Cmanifold Mc of the entire robot into R 9? If so, we would achieve a significant model reduction. Let zn|Mn and z|Mc be, respectively, a point on the last link C-manifold Mn sitting in R 9 and a point on the entire robot C-manifold Mc in R 9n. If we define a common local coordinate system by using the n robotic joint positions q = (q1 \u00b7 \u00b7 \u00b7 qn)T on both Mn and Mc, then we seek a continuous one-to-one in both directions between Mn and Mc via the common local coordinates q, i.e., zn|Mn \u2190\u2192 q|Z \u2190\u2192 z|Mc . We can directly observe that through the common q under the quotient operation by the integer group Z for every joint angle, each point on Mn determined by zn = \u03b6n(q) = Zn 0 (q)\u03ben can respond to a unique point on Mc determined by z = \u03b6(q) = Z(q)\u03be, and vice verso, as long as the mapping \u03b6n : Mn \u2192 R 9 of the last link is a smooth embedding under the robotic joint positions in q as the local coordinate system. This reaches the following lemma: Lemma 3. For an open serial-chain robotic system with n joints, let the n joint positions in q = (q1 \u00b7 \u00b7 \u00b7 qn)T be defined as a common n-tuple local coordinate system on both the C-manifold Mn of the last link and C-manifold Mc of the entire robot. If (7.54) with i = n can be a smooth embedding that sends Mn into R 9, then Mn is diffeomorphic to Mc, i.e., Mn Mc. Note that if the number of joints of the robot arm is n \u2264 6 in the above lemma, it may have more chances for the mapping \u03b6n to be a smooth embedding. If n > 6, such an embedding condition on Mn would not be possible. In fact, n > 6 is a redundant robot case, as discussed in Chapter 5, and every redundant robotic system has a continuum of multi-configurations. In the redundant robot case, the combined C-manifold Mc of the entire robot may still be 6-dimensional and so is the C-manifold Mn of the last link, unless additional subtasks are augmented to the 6 d.o.f. rigid motion as a redundant kinematic output. In addition, Lemma 3 reveals a differentiable (smooth) topological equivalence between the two C-manifolds Mn Mc under the condition that (7.54) with i = n is a smooth embedding if all the joint positions of the robot constitute a local coordinate system. A number of non-redundant robot arms can directly satisfy the above condition due to their special mechanical structures. A robot having more independent prismatic joints may have more chances to satisfy the smooth embedding condition. If, in general, the smooth embedding condition cannot hold by the last link mapping alone, we have to augment the Euclidean zspace by adding one or more links until the augmented mapping is qualified to be a smooth embedding. Obviously, if the final augmentation covers all n links of the robot, it must be smoothly embeddable, because the total Riemannian metric W = JT J that is also the inertial matrix of the robot should always be non-singular. However, in most cases, only a small number of links need to be selected, say 1 < k < n, which, of course, include the last link to form an augmented mapping into R 9k to meet the smooth embeddability condition. Such an augmented smooth embedding \u03b6nk with the smallest k is referred to as the minimum embedding of the C-manifoldMc for the robotic system. Therefore, based on the principle of Lemma 3, under the smallest number k, it is always true that Mnk Mc for every open serial-chain robotic system, even for a redundant robot. Now, the new question is how to find such a minimum embedding? Recalling Definition 6, a mapping between two manifolds can be an embedding if it is a one-to-one immersion. Thus, by intuition, this one-to-one condition holds if no more multi-configuration case can be found among the remaining n\u2212 k links (k < n) whenever the selected k links are stationary. It is well known in robotic kinematics that a 6-joint PUMA robot with a shoulder offset (i.e., d2 = 0) has up to eight different configurations (i.e., eight different sets of I-K solutions) if the last link (link 6) is fixed. Whereas for the 6-joint PUMA-like arm without the shoulder offset (d2 = 0), the number of different configurations is down to four. In contrast, a 6-joint Stanford robot with a shoulder offset (d2 = 0) has in total four different configurations whenever link 6 is motionless. These four different sets of the I-K solutions for the Stanford arm have been animated in MATLABTM , as realistically depicted in Figures 7.14 and 7.15. Therefore, to only pick the last link of the 6-joint Stanford arm (or PUMA arm) is insufficient to constitute a smooth embedding for the C-manifold Mc if all the 6 joint positions are defined as the local coordinates on the last link C-manifold M6. However, it is also observable that if we further select link 1 and link 4, in addition to link 6 for the Stanford arm to augment a relatively larger mapping for Mc, then it becomes embeddable because all the possible multi-configurations disappear whenever the selected three links are stationary. In this case, the minimum embedding \u03b6nk is determined by augmenting the mappings (7.54) of link 6, link 4 and link 1, and thus k = 3. The total dimension of this minimum embedding is now 9k = 9 \u00d7 3 = 27 that is much less than 9n = 9\u00d7 6 = 54 in the regular augmentation (7.55), and just touches the dimension predicted by the Nash-Greene Theorem for isometrically embedding the C-manifold Mc of the Stanford arm. Furthermore, if each link can adopt the two-point model, i.e., only one radial vector r1c for each link with a diagonal inertia tensor, then the dimension of the minimum embedding will be further reduced to 6k = 18. Actually, to test whether a mapping \u03b6 : Mn \u2192 R m for an n-dimensional C-manifold Mn is an embedding, we may just check its Jacobian matrix J = \u2202\u03b6/\u2202q to see if rank(J) = n, the full rank at every point on Mn. If the Jacobian matrix of a mapping has full rank, then every multi-configuration case will disappear. Therefore, to find the minimum embedding of the Cmanifold Mc for a robot by the above k-augmentation procedure, testing its Jacobian matrix J to see if it is full-ranked is the easiest way to forecast the result. Moreover, two smoothly topologically equivalent (diffeomorphic) Riemannian manifolds in a common ambient Euclidean space can always be smoothly deformed to make them isometric to each other. Even though their ambient Euclidean spaces are in different dimensions, it may still be possible, but not always. We refer to this smooth deformation process as an isometrization [23]. Since the Riemannian metric of any smooth C-manifold is given by W = JTJ , where J is the Jacobian matrix of the C-manifold, based on (7.55) and (7.56), each component of W can be written as wi j = gTi gj = \u03beT \u2202Z \u2202qi T \u2202Z \u2202qj \u03be = tr ( \u2202ZT \u2202qi \u2202Z \u2202qj \u03a8 ) , for i, j = 1, \u00b7 \u00b7 \u00b7 , n, (7.57) where \u03a8 = \u03be\u03beT is called a parameter matrix and tr(\u00b7) is the trace of a square matrix (\u00b7). We now give a formal definition regarding the isometrization below: Definition 8. Let two n-dimensional C-manifolds Mn a and Mn b be smoothly embedded into R ma and R mb , respectively, on both of which a common local coordinate system {q1, \u00b7 \u00b7 \u00b7 , qn} is defined. Then,Mn a is said to be isometrizable to Mn b if for each element (wa) i j of the metric Wa on Mn a , i.e., (wa) i j = \u03beTa \u2202ZT a \u2202qi \u2202Za \u2202qj \u03bea = tr ( \u2202ZT a \u2202qi \u2202Za \u2202qj \u03a8a ) , there is a real parameter function D(\u00b7) such that the new parameter matrix \u03a8 \u2032 a = D(\u03a8b) can make the following equation hold globally: (w\u2032 a) i j = tr ( \u2202ZT a \u2202qi \u2202Za \u2202qj \u03a8 \u2032 a ) = tr ( \u2202ZT b \u2202qi \u2202Zb \u2202qj \u03a8b ) = (wb) i j for each i, j = 1, \u00b7 \u00b7 \u00b7 , n. We call this real parameter function D(\u00b7) a deformer. Therefore, to respond to the model reduction challenge, the most promising approach is to expect the minimum embedding \u03b6nk of the C-manifold Mnk to be isometrizable to the combined C-manifold Mc through the above deformation with a certain deformer. However, Mc has been isometrically embedded into R 9n, and may possess more geometrical details than Mnk that is embedded into a smaller Euclidean space R9k with k < n. In other words, R9k may not be spatial enough to allowMnk to be isometrizable toMc even though topologically Mnk Mc. Nevertheless, we are dealing with a special open serial-chain robot case. As commonly experienced, when trying to derive a symbolical form of kinetic energyKi for link i of an n-joint serial-chain robotic system with (1 < i \u2264 n), we must take into account every velocity of both translation and rotation of link i \u2212 1, in addition to all the terms contributed by the motion of link i itself, because the latter is often imposed by the former. Therefore, the sum Knk of all the kinetic energies Ki\u2019s for the k links involved in the C-manifold Mnk that is associated with the minimum embedding \u03b6nk should be able to cover every factor contained in the total kinematic energy K = \u2211n i=1Ki of the robot. In other words, we can always adjust each dynamic parameter in Knk to make Knk equal to K. Because the coefficient of each q\u0307iq\u0307j in K given by (7.47) is just the metric element wij , this parameter adjustment is virtually a smooth deformation mentioned in Definition 8. Hence, we reach the following theorem: Theorem 4. For an n-joint open serial-chain robot dynamic system, its minimum embeddable C-manifold Mnk is diffeomorphic to the combined Cmanifold Mc of the entire robotic system. If all the robotic joints are revolute, then, Mnk is also isometrizable to Mc. This theorem underlies a fundamental principle of robot dynamic model reduction. Namely, the lower-bound of dynamic model reduction in the sense of topology is a subsystem with the minimum embeddable C-manifold for every open serial-chain robotic system. In other words, the robot dynamic model cannot be further reduced below the lower-bound, otherwise both the embeddability and isometrizability will no longer be guaranteed, and may even cause a catastrophe of topological structure destruction [38, 39]. Theorem 4 indicates that if every joint is revolute for a robotic system, then the minimum embeddable C-manifold Mnk is not only diffeomorphic to the combined C-manifold Mc, but is also isometrizable to Mc. If one of the joints is prismatic, Mnk is still diffeomorphic to Mc, but Mnk may have to further include the prismatic link in order to meet the isometrizability. The reason is that since the variable for a prismatic joint is a length di, instead of an angle, as one of the local coordinates, it will often be mixed with the dynamic parameters to be deformed together in the Riemannian metrics {wij}, causing unavailability of an effective deformer to distinguish the joint variable di from the dynamic parameters. As a first example, let us look at a well-known inverted pendulum system that consists of only two links: a linearly moving cart and a rotating pole mounted on the cart, as shown in Figure 7.16. With the joint positions x1 and \u03b82 defined as two local coordinates q = (x1 \u03b82) T for its 2-dimensional combined C-manifold, we can readily find its inertial matrixW by extracting all the coefficients of q\u0307iq\u0307j terms for i, j = 1, 2 from its kinetic energy K, i.e., W = ( m1 +m2 \u2212m2lc2s2 \u2212m2lc2s2 m2l 2 c2 + I2 ) , (7.58) where m1 and m2 are the masses of the cart and the top pole as link 1 and link 2, respectively, lc2 is the length between the revolute joint axis and the mass center of the pole, as link 2, and I2 is the moment of inertia of link 2. It can be directly seen that if link 2 is fixed, the cart (link 1) will have no chance to move. Thus, link 2 determines the minimum embeddable Cmanifold Mnk of the system with n = 2 and k = 1. For this planar system, the position vector of the mass center of link 2 with respect to the base is pc20 = ( ax1 + bc2 bs2 ) , where a and b are two dynamic parameters of link 2. Since this pole rotates only about one single axis, we can simply use h\u03b82 with a new parameter h to represent the rotation of link 2 under 0 \u2264 \u03b82 < 2\u03c0. Therefore, the smooth embedding for Mnk can be written as \u23a7\u23a8 \u23a9 z12 = ax1 + bc2 z22 = bs2 z32 = h\u03b82. (7.59) The Jacobian matrix of this minimum embedding z2 = \u03b62(q) becomes J2 = \u2202\u03b6 \u2202q = \u239b \u239d a \u2212bs2 0 bc2 0 h \u239e \u23a0 , the rank of which is obviously rank(J) = 2, as long as both a = 0 and h = 0. Its Riemannian metric turns out to be W2 = JT 2 J2 = ( a2 \u2212abs2 \u2212abs2 b2 + h2 ) . Comparing this with the inertial matrix W in (7.58), we can see that if a = \u221a m1 +m2, b = m2lc2\u221a m1 +m2 and h = \u221a m1m2 m1 +m2 l2c2 + I2, (7.60) thenW2 =W . This demonstrates that the minimum embeddable C-manifold given by (7.59) is isometrizable to the combined C-manifold Mc of the inverted pendulum system, and equation (7.60) is just its deformerD(\u03a8) during the isometrization process. Figure 7.17 visualizes the minimum embeddable C-manifold M21 in R 3 based on (7.59) when a = 4, b = 1.5 and h = 1 in MATLABTM . This 2- surface is diffeomorphic to the compact cylindrical surface S1 \u00d7 I1 [28, 29] after \u03b82 in z32 is confined within [0, 2\u03c0) by a quotient operation R 1/Z, where Z is the integer additive group. In other words, once the z32 component of the 2-surface reaches 2\u03c0, it should be imagined to return immediately back to the zero and start over again. The second example is a three-revolute-joint (RRR-type) planar arm, as shown in Figure 7.18. According to equation (7.54), the last link i = n = 3 has a 3-dimensional C-manifold M3 that can be sent to R 3 by \u23a7\u23a8 \u23a9 z13 = ac1 + bc12 + dc123 z23 = as1 + bs12 + ds123 z33 = h(\u03b81 + \u03b82 + \u03b83) (7.61) with four non-zero parameters a, b, d and h, where sij = sin(\u03b8i + \u03b8j), cij = cos(\u03b8i + \u03b8j), sijk = sin(\u03b8i + \u03b8j + \u03b8k) and cijk = cos(\u03b8i + \u03b8j + \u03b8k) for i, j, k = 1, 2, 3. Its Jacobian matrix becomes J3 = \u2202\u03b63 \u2202q = \u239b \u239d \u2212as1 \u2212 bs12 \u2212 ds123 \u2212bs12 \u2212 ds123 \u2212ds123 ac1 + bc12 + dc123 bc12 + dc123 dc123 h h h \u239e \u23a0 . (7.62) The determinant of this 3 by 3 Jacobian matrix is calculated as det J3 = abhs2 = abh sin \u03b82. This shows that M3 is singular at \u03b82 = 0 or \u00b1\u03c0. In fact, we can see that for each fixed z3 = (z13 z 2 3 z 3 3) T \u2208 R 3, \u03b81 through \u03b83 can have two different sets of I-K solution, as long as they keep the sum \u03b81 + \u03b82 + \u03b83 fixed while one uses \u03b82 and the other one takes \u2212\u03b82, as shown in Figure 7.18. This is a typical multi-configuration phenomenon. We may thus construct a minimum embeddable C-manifold by adding link 1, which will clearly make the above multi-configuration disappear. Since link 1 has only a single-axis rotation, we can augment (7.61) by z4 = h1\u03b81 so that a possible minimum embedding would be \u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 z1 = ac1 + bc12 + dc123 z2 = as1 + bs12 + ds123 z3 = h(\u03b81 + \u03b82 + \u03b83) z4 = h1\u03b81 (7.63) with a new non-zero parameter h1. It can be easily verified that the rank of the following new 4 by 3 Jacobian matrix J is always full: J = \u2202\u03b632 \u2202q = \u239b \u239c\u239d \u2212as1 \u2212 bs12 \u2212 ds123 \u2212bs12 \u2212 ds123 \u2212ds123 ac1 + bc12 + dc123 bc12 + dc123 dc123 h h h h1 0 0 \u239e \u239f\u23a0 , and the set of all four equations in (7.63) is also one-to-one. Thus, it shows that equation (7.63) is a minimum embedding with n = 3 and k = 2. The Riemannian metricW32 on the minimum embeddable C-manifold can be found by W32 = JTJ = \u239b \u239d w11 w12 w13 w12 b2 + 2bdc3 + d2 + h2 bdc3 + d2 + h2 w13 bdc3 + d2 + h2 d2 + h2 \u239e \u23a0 , where w11 = a2 + b2 + 2abc2 + 2adc23 + 2bdc3 + d2 + h2 + h21, and w12 = b2 + abc2 + adc23 + 2bdc3 + d2 + h2, w13 = adc23 + bdc3 + d2 + h2. One can also verify without difficulty that by adjusting all the parameters a, b, d, h and h1, the metric W32 = JTJ can be equal to the inertial matrix W of this RRR-type planar robot. Such a parameter adjustment just plays the role of deformer D(\u03a8) in the C-manifold isometrization. In contrast to the above RRR-type planar robot arm, let us revisit the RPR-type planar robot, as given in Figure 7.7 or 7.9. At first glance, it seems sufficient enough to pick up only the last link to represent the minimum embeddable C-manifold for this particular robot, because the configuration becomes unique whenever the last link is motionless. If so in this case, n = 3 and k = 1. Based on the previous kinematics analysis of this planar robot, the tip position vector and the position vector of frame 2 can be found as pt0 = \u239b \u239d d2s1 + d4s13 \u2212d2c1 \u2212 d4c13 0 \u239e \u23a0 , and p20 = \u239b \u239d d2s1 \u2212d2c1 0 \u239e \u23a0 where d4 is the length of the last link. Since the z-component of pt0 is zero due to the planar arm, while the number of joints is three, we must replace this zero by the orientation of the last link, which can be uniquely determined by \u03b81 + \u03b83. Thus, the proposed embedding is given as follows: \u23a7\u23a8 \u23a9 z13 = ad2s1 + bs13 z23 = \u2212ad2c1 \u2212 bc13 z33 = h(\u03b81 + \u03b83) (7.64) with parameters a, b and h to be adjusted. Then, the Jacobian matrix becomes J3 = \u2202\u03b63 \u2202q = \u239b \u239d ad2c1 + bc13 as1 bc13 ad2s1 + bs13 \u2212ac1 bs13 h 0 h \u239e \u23a0 . Its determinant is det(J) = \u2212a2hd2 and is never zero as long as the prismatic joint value d2 = 0. Now, the Riemannian metric for this minimum embeddable C-manifold M31 turns out to be W31 = JT 3 J3 = \u239b \u239d a2d2 + b2 + h2 + 2abd2c3 \u2212abs3 abd2c3 + b2 + h2 \u2212abs3 a2 \u2212abs3 abd2c3 + b2 + h2 \u2212abs3 b2 + h2 \u239e \u23a0 . Comparing this with the inertial matrixW of this RPR arm derived in equation (7.23), we find that if a2 = m2 +m3, ab = m3lc3, and b2 + h2 = I3 +m3l 2 c3, every element wij in W can be matched by W31 except the first one w11 that contains a mixed term m2(d2 \u2212 lc2)2 between the prismatic joint variable d2 and the dynamic parameters. Therefore, while equation (7.64) is an embedding, it has not been isometrizable yet to exactly match the inertial matrix W . This phenomenon reveals a special complication due to the prismatic joint in isometrization. By further augmenting the second prismatic link, the new embedding becomes: \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 z13 = a1d2s1 + b1s13 z23 = \u2212a1d2c1 \u2212 b1c13 z33 = h1(\u03b81 + \u03b83) z43 = a2(d2 \u2212 b2)s1 z53 = \u2212a2(d2 \u2212 b2)c1 z63 = h2\u03b81 (7.65) along with six parameters a1, b1, h1, a2, b2 and h2 to be adjusted, where d2 \u2212 b2 indicates a distance to converge to the mass center of link 2. The Jacobian matrix of the new proposed embedding (7.65) can be derived as J3 = \u2202\u03b63 \u2202q = \u239b \u239c\u239c\u239c\u239c\u239c\u239d a1d2c1 + b1c13 a1s1 b1c13 a1d2s1 + b1s13 \u2212a1c1 b1s13 h1 0 h1 a2(d2 \u2212 b2)c1 a2s1 0 a2(d2 \u2212 b2)s1 \u2212a2c1 0 h2 0 0 \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 . Thus, the Riemannian metric becomes W32 = JT 3 J3 = \u239b \u239d w11 \u2212a1b1s3 a1b1d2c3 + b21 + h21 \u2212a1b1s3 a21 + a22 \u2212a1b1s3 a1b1d2c3 + b21 + h21 \u2212a1b1s3 b21 + h21 \u239e \u23a0 , where w11 = a21d 2 2 + b21 + h21 + 2a1b1d2c3 + a22(d2 \u2212 b2)2 + h22. In comparison with equation (7.23) again, the new Riemannian metricW32 has now sufficient factors to match the inertial matrix W . Namely, let a1 = \u221a m3, a2 = \u221a m2, b1 = \u221a m3lc3, h1 = \u221a I3, and b2 = lc2, h2 = \u221a I1 + I2 +m1l2c1. Then every element wij inside the new metric W32 can exactly match with the inertial matrix W in (7.23). The last example is to dynamically model a general fully parallel-chain mechanism, such as a 6-6 Stewart platform, as studied in [36, 37]. If we focus on the top mobile disc, according to equation (7.25), its kinetic energy can be written as follows: K = 1 2 mv6T0 v60 +mv6T0 CT 6 \u03c9 6 0 + 1 2 \u03c96T 0 \u03936\u03c9 6 0, where m and \u03936 are the mass and inertia tensor with respect to frame 6 on the top disc, respectively, and C6 is the skew-symmetric matrix of the mass center coordinates with respect to frame 6. If the origin of frame 6 is defined at the mass center of the top mobile disc, the second term of the kinetic energy K vanishes and \u03936 = \u0393c. Because the top disc is a single rigid body, Theorem 3 can be directly applied to it. Then, an isometric embedding that can send the 6-dimensional C-manifold M6 6 of the top disc motion to Euclidean 9-space R 9 is given by z = \u03b6(q) = Z\u03be = \u239b \u239d R6 0 p60 O R6 0 O R6 0 \u239e \u23a0 \u03be (7.66) with a 10 by 1 dynamic parameter vector \u03be based on equation (7.54). In order to find the Riemannian metric W = JT J endowed on the CmanifoldM6 for the top mobile disc of the 6-6 Stewart platform, it is required to know a local coordinate system q to be defined on M6. If we pick the 6 prismatic joint lengths qi = li\u2019s to form the local coordinate system, then intuitively, the top disc will have no multi-configuration chance so that the mapping (7.66) is a minimum embedding for the C-manifold Mc of the entire Stewart platform. The Jacobian matrix of the minimum embedding should be J = \u2202\u03b6/\u2202q. However, this turns back to the forward kinematics (F-K) problem. Namely, it is difficult or even impossible to find an explicit closed form for either R6 0 or p60 as a function of q = (q1 \u00b7 \u00b7 \u00b7 q6)T = (l1 \u00b7 \u00b7 \u00b7 l6)T before taking derivatives to determine J and then W . If we define a local coordinate system alternatively other than the 6 piston lengths, it may be possible to reach a Riemannian metric result. For instance, let the 6 local coordinates be defined by the variables in Cartesian space: qc = \u239b \u239c\u239c\u239c\u239c\u239c\u239d x y z \u03c6 \u03b8 \u03c8 \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 = ( p60 \u03c160 ) , (7.67) where x, y and z are the three coordinate components of p60, and \u03c6, \u03b8 and \u03c8 are the roll, pitch and yaw Euler angles to represent R6 0, see equation (3.7) in Chapter 3. Namely, p60 = \u239b \u239d x y z \u239e \u23a0 , R6 0 = \u239b \u239d c\u03c6c\u03b8 \u2212s\u03c6c\u03c8 + c\u03c6s\u03b8s\u03c8 s\u03c6s\u03c8 + c\u03c6s\u03b8c\u03c8 s\u03c6c\u03b8 c\u03c6c\u03c8 + s\u03c6s\u03b8s\u03c8 \u2212c\u03c6s\u03c8 + s\u03c6s\u03b8c\u03c8 \u2212s\u03b8 c\u03b8s\u03c8 c\u03b8c\u03c8 \u239e \u23a0 . Then, the 9 by 1 column of the embedding z = \u03b6(qc) = Z\u03be in (7.66) after the structure matrix Z is linearly combined by the dynamic parameters in \u03be can be expressed explicitly in terms of the above 6 new local coordinates. The Jacobian matrix J = \u2202\u03b6/\u2202qc as well as the Riemannian metric W = JT J can also be determined successfully. However, under the new local coordinate system (7.67), due to the nonuniqueness between the roll, pitch and yaw Euler angles and the rotation matrix, see the example in Chapter 3, there will be at least one multiconfiguration case. In other words, the mapping z = \u03b6(qc) will no longer be a minimum embedding. Even if it could be remedied by imposing a constraint on the three Euler angles, once the Riemannian metric W on the Cmanifold is determined, by substituting it into the Lagrange equation (7.15), the control input \u03c4c will no longer be the desired six piston forces. Instead, it should be a some Cartesian force vector in correspondence to the new local coordinates defined in Cartesian space and given by (7.67), which requires a conversion. Naturally, we may take advantage of the statics equation (5.30) from Chapter 5 for the Stewart platform to convert the control input \u03c4c that is resolved by the Lagrange equation in Cartesian space to a piston joint force vector by \u03c4 = J\u22121 0 F0. To do so, according to equation (3.12) in Chapter 3, the angular velocity of the top disc is \u03c96 0 = \u239b \u239d 0 \u2212s\u03c6 c\u03c6c\u03b8 0 c\u03c6 s\u03c6c\u03b8 1 0 \u2212s\u03b8 \u239e \u23a0 \u239b \u239d \u03c6\u0307 \u03b8\u0307 \u03c8\u0307 \u239e \u23a0 = D\u03c1\u030760. Hence, the 6 by 1 Cartesian velocity vector becomes V0 = ( v60 \u03c96 0 ) = ( I O O D )( p\u030760 \u03c1\u030760 ) = Bq\u0307c, where the 6 by 6 coefficient matrix is denoted by B, and I and O are the 3 by 3 identity and zero matrix, respectively. Based on the Jacobian equation (5.29) in Chapter 5 for the Stewart platform, q\u0307 = JT 0 V0 = JT 0 Bq\u0307 c. (7.68) This arrives at a kinematic conversion in tangent space between the prismatic joint position vector q = (l1 \u00b7 \u00b7 \u00b7 l6)T and the local coordinate system qc defined by (7.67). Furthermore, the mechanical power seen in joint space is P = q\u0307T \u03c4 , where \u03c4 = (f1 \u00b7 \u00b7 \u00b7 f6)T is the piston actuating force vector, while the same power seen in Cartesian space is P = q\u0307cT \u03c4c for a Cartesian force \u03c4c that is corresponding to the local coordinates in (7.67). Based on the principle of energy conservation, q\u0307T \u03c4 = q\u0307cT \u03c4c. Substituting (7.68) into here, we have q\u0307cTBTJ0\u03c4 = q\u0307cT \u03c4c so that \u03c4c = BTJ0 \u03c4. (7.69) This new statics equation is similar to (5.30) in Chapter 5. Thus, if a control law \u03c4c can be resolved through the Lagrange equation (7.15), then the piston actuating forces can be determined by the new statics equation (7.69), i.e., \u03c4 = J\u22121 0 B\u2212T \u03c4c, provided that both the Jacobian matrix J0 and the matrixB are non-singular. Therefore, the conversion issue between \u03c4c and \u03c4 has been solved by the statics in (7.69), but the one-to-one, or multi-configuration issue still remains unfixed as a barrier of achieving the minimum isometrizable embedding. In conclusion, if we insist in using the 6 piston lengths to define a local coordinate system on the C-manifoldMc of the Stewart platform, we will not be able to continue its Riemannian metric determination towards a successful adaptive control design until the F-K problem for the closed parallel-chain systems is resolved. Finally, it should be pointed out that all the parameter adjustments in the above examples are just to illustrate the possibility of isometrization, and are not required to do so by hand at all. Instead, the computer program will automatically adjust the model parameters towards the isometrization if an effective adaptive control algorithm is implemented. The detailed introduction and discussion on adaptive control as well as a 3D robotic system example will be given in the next chapter."
        ]
    },
    {
        "image_filename": "designv11_100_0002724_s1052618811020142-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002724_s1052618811020142-Figure3-1.png",
        "caption": "Fig. 3. The primary deviations of the table (slide).",
        "texts": [
            " If the drive operating according to the X coordinate was ideal, the table was directly in the rated position xrated. If the drive is not ideal, the table is under the real position xreal relative to the X axis with the deviation of \u03b4x(x) = xreal \u2013 xrated. At the same time, the table forms the sliding pair jointly with the column, in which the real conjugate surfaces are characterized by the deviations from the ideal ones (rated). Respectively, in the general case, the table (as a solid body in space) is characterized by the deviations with respect to five other coordinates (Y, Z, A, B, and C). Figure 3 depicts the primary deviations of the table (slide), which is moved according to the straightline guides (linear coordinate x) [6] and according to ISO 230 1: the linear deviations (\u03b4x, \u03b4y, and \u03b4z) along X, Y, and Z axes and the angular deviations (\u03b1x, \u03b1y, and \u03b1z) (rotations around X, Y, and Z). These deviations are primary deviations of the slide link (1). To build a model that generates the integral deviation according to primary deviations, it is convenient to use the equations for deviations \u03b4x(xrated), \u03b4y(xrated), \u03b4z(xrated), \u03b1x(xrated), \u03b1y(xrated), and \u03b1z(xrated) written with respect to the coordinate xrated of the undisturbed (rated) translatory motion of the relatively immov able slide coordinate system, as related to the column",
            " 315] the solid can slide from one state into another conjugate one via dif ferent translatory displacements that depend on how we choose the point (pole), whose position deter mines the translatory motion. From here, it follows that when we measure \u03b4x(xrated), \u03b4y(xrated), \u03b4z(xrated), \u03b1x(xrated), \u03b1y(xrated), and \u03b1z(xrated) deviations, it is necessary to take into account that they are related to a JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 40 No. 2 2011 PRIMARY DEVIATIONS OF MECHANISM LINKS WITH SLIDING 109 certain pole and all measurements should be performed relative to this pole P (Fig. 3). Pole P is placed into the point of intersection of the measurement lines. The origin O1 of coordinate system X1Y1Z1 is placed into the same point. If we choose another pole, the measurement results should be recalculated considering the coordinates of old and new poles and of axes of rotation. Hereby, the initial information needed for generating the model that describes the integral deviation of the table (slide 1) at the guides (sliding pair 6) is as follows (Fig. 3): \u2013 \u03b4x(xrated) = xreal \u2013 xrated deviation during positioning along X coordinate (EXX, line 5); \u2013 deviation from linearity \u03b4y(xrated) that takes place in XY plane (EYX, line 7); \u2013 deviation from linearity \u03b4z(xrated) that takes place in XZ plane (EZX, line 3); \u2013 \u03b1x(xrated) rotation around X axis (EAX, line 4); \u2013 \u03b1y(xrated) rotation around Y axis (EBX, line 8); \u2013 \u03b1z(xrated) rotation around Z axis (ECX, line 2); \u2013 coordinates of measurement lines of \u03b4x(xrated), \u03b4y(xrated), and \u03b4z(xrated) linear deviations by assuming that the pole of motion is in the point of their intersection"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.6-1.png",
        "caption": "Fig. 3.6 Side view of a bifilar pendulum in the rest position",
        "texts": [
            "8) More generally, combinations of the two modes of oscillation are possible and the point mass, m2, moves on the surface of a torus generated by the rotation of a circle with centre at the centre of the Y and radius l2 about the line joining the clamps. In a bifilar pendulum (Gauld 2005) a point mass, m, is suspended from a pair of inextensible massless strings of equal length arranged so that the path of the point mass, m, is constrained to a vertical plane. A side view of a bifilar pendulum in the rest position shown in Fig. 3.6. The behaviour of a bifilar pendulum is the same as that of a simple rod pendulum (Sect. 2.3), provided that the strings remain taut. Hence, if the resolved length of the strings in a vertical direction is l , then for small amplitudes the pendulum period, P , given approximately by Eq. 2.15. Fig. 3.7 Rotating simple rod pendulum. If the angular velocity, !, is constant it is a circular rod pendulum The rotating simple rod pendulum (Fig. 3.7), is a simple rod pendulum (Fig. 2.1a) with the frictionless pivot on a rigid massless vertical axle free to rotate in frictionless bearings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003465_ijrapidm.2011.040688-Figure17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003465_ijrapidm.2011.040688-Figure17-1.png",
        "caption": "Figure 17 FDM tensile specimen nested within build volume (see online version for colours)",
        "texts": [
            " Known in industry as design allowables, these data offer substantiation for candidate parts to perform in relative operational conditions. The larger the database, the more credibility is added for AM to act as a legitimate manufacturing process. Finite element analysis packages generally allow for isotropic property input for modulus; therefore, the design engineers will use Z-axis modulus data to determine a worst-case condition for operating usage environments. Next, candidate parts are selected to exploit the FDM process. The tensile specimen is placed within the build volume as shown in Figure 17. Each tensile bar offers a unique build pattern to match closely to a part fabricated in the FDM process. This build pattern allows for part geometry to be constructed with most feature definition simulated via tensile bar grouping with the test geometry shown in Figure 17. Once a representative part is constructed within a build volume, as depicted in Figure 17, the tensile specimen is removed from the build chamber, broken into individual test specimens and tensile tested per ASTM D638 conditioning. Data are analysed to provide a clear understanding of the range of raster orientations contribution to Z-axis tensile strength. Part candidate selection goes hand-in-hand with mechanical property evaluation. These parts are generally highly complex, low quantity geometries that may be polymer based. These parts are typically difficult to construct using conventional manufacturing solutions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002978_detc2011-48760-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002978_detc2011-48760-Figure3-1.png",
        "caption": "Figure 3. Meshing of pinion and gear.",
        "texts": [
            " When the tooth surface of the complementary crown gear in O-XYZ is transformed into the coordinate system fixed to the generated gear, the tooth surface of the straight bevel gear is expressed. The left and right tooth surfaces of the pinion are expressed as Xp and Xp', respectively. The right and left tooth surfaces of the gear are expressed as Xg and Xg', respectively. Moreover, the unit normals of Xp, Xp' Xg, and Xg' are expressed as Np, Np' Ng, and Ng'. Henceforth, the subscripts \u201cp\u201d and \u201cg\u201d indicate that each is related to the pinion and gear, respectively. The generated pinion and gear are assembled in a coordinate system Oh-xhyhzh as shown in Fig. 3. Suppose that p and g are the rotation angles of the pinion and gear, respectively. The position vectors of the pinion and gear tooth surfaces must coincide and the direction of two unit normals at this position must be also coincide in order to contact the two surfaces. Therefore, the following equations yield: ),()(),()( ),()(),()( ggggpppp ggggpppp uu uu   NCNB XCXB   (11) where B and C are the coordinate transformation matrices for Downloaded From: http://proceedings.asmedigitalcollection.asme"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002992_scis-isis.2012.6505132-Figure18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002992_scis-isis.2012.6505132-Figure18-1.png",
        "caption": "Fig. 18. Outside appearance of simulation",
        "texts": [
            " \u03c9(q(t)) = \u221a detJn(q(t))J T n (q(t)) (14) Observed Fig. 11, Fig. 13, Fig. 15 and Fig. 17, we obviously can believe that predictive control can also predict the manipulability degree of manipulator. However, in Fig. 11, when t=9 the value of manipulability degree get large suddenly, and manipulability degree become difficult to be predicted. About this problem, we also need to do further study. We want to know the manipulator will be predicted effectively or not when the trajectory is curve trajectory, The solid line in Fig. 18 expresses a target curve trajectory set to be followed. The angle of actual manipulators\u2019 link 1 and the predictive angles q\u03021(t \u2217 1), q\u03021(t \u2217 2), q\u03021(t \u2217 3) of manipulators\u2019 link 1 are respectively indicated in Fig.19 and Fig. 21 when predictive interval time are 0.3[s] and 0.15[s]. The angle of actual manipulators\u2019 link 2 and the predictive angles q\u03022(t \u2217 1), 978-1-4673-2743-5/12/$31.00 \u00a92012 IEEE 206 q\u03022(t \u2217 2), q\u03022(t \u2217 3) of it are respectively indicated in Fig. 20 and Fig. 22 when predictive interval time are 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002506_0029-554x(59)90172-7-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002506_0029-554x(59)90172-7-Figure1-1.png",
        "caption": "Fig. 1. Corner region of a rectangular core.",
        "texts": [
            " Fagg, Error at a Rectangular Corner in Relaxation Methods, Nature 182 (1958) 600. Butler and Fagg4). They solved the boundary value problem for ~o using relaxation methods. In this paper, we have used numerical methods to solve the general square core problem both with and without an air gap. A complete description of the numerical methods used will be given in section 4 and appendix 1. Before proceeding with this, let us consider the idealized case of the field near a square corner when all other boundaries are removed to infinity (fig. 1). By means of a Schwarz-Christoffel transformation it can be shown that the flux function ~0 is given by w ~ \u00a2 + ito ~ o: (Re~O)2/3 (1) where c~ is a constant. The lines of force are given by ~o = const, and the two field components are given by 8%e 1 0~v H o = ~ a n d H r . . . . R O 0 \" From equation (1), z? = \u2022R 2/3 sin ~ 0 H o = {aR-1 /3 sin { 0 (2) H r = - - ~ a R -1/3cos ~0 and H = %/Ho2 + H r 2 = ~a \u00b0:R-l~3. the eddy current loss Pc, is given by P c = Ke J~Rlfo~=H~R dOdR = \u00bd Ke\u00b0:~zR14/3 (3) where Ke is a constant depending on the resistivity of the material",
            " (4) ~zR12 P h = K h f o R I ~ ~ I - I a R d O d R = ~4K (5) (Ph)ave l~ Kh c\u00a2a Rl -1 (6) = 27 Eqs. (3) and (5) show that the losses are finite for finite values of R~. However, the average loss M A G N E T I C L O S S E S IN C O R E S OF V A R I O U S S H A P E S 135 per uni t volume, given by eqs. (4) and (6), increases wi thout bound as R 1 tends to zero. There is clearly a very large loss concentrat ion at the inside corners of a square core. The solution is exact when the outs ide boundary of the core coincides with a line of force such as AB in fig. 1. We can obtain an approximat ion to the losses for a square core by integrat ing the flux function ~p given by eq. (2) over a region bounded by CDE as the outer boundary. This will not be exact, since the line CDE is not a line of force. Then P h = K h ] ] H 8 d x d y ~ 1 . 5 6 7 K h a 8b C D E F O G where b is the width of the core. Choosing v 2 -= 0 on boundary GOF and 7J = 1 on CDE, a can be evaluated by using the condition ~ = 1 when R = b and 0 = \u00bdz~. This choice of a point on the outer boundary is quite arbi t rary but it does give an approximate value of ~"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002215_978-3-0348-0130-0_35-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002215_978-3-0348-0130-0_35-Figure2-1.png",
        "caption": "Fig. 2 Movement direction of robot ri using Move-On-Bisector",
        "texts": [
            " Corollary 1 (See [4]) When the continuous variant of the Go-To-The-Middle strategy is performed, the maximum distance travelled by a relay is (n) for a worst-case start configuration. A more natural strategy in this setting is the Move-On-Bisector strategy. In this strategy, a robot ri moves in direction of the angle bisector of the angle formed at ri \u2019s position between the directions in which the two neighbours ri\u22121 and ri+1 are positioned. This angle is called \u03b1i(t) when measured at time t . See Fig. 2 for a visualisation. When using this strategy, the robots do not only converge to their destination points, but they actually reach them within finite time. We will see that the maximum distance travelled by the robots is also (n) in the worst case, but we can analyse this strategy even in more detail: we do not only analyse worst case instances, but we will compare the maximum distance travelled when using Move-On-Bisector for each individual start configuration with the maximum distance OPT travelled when using an optimal global algorithm for the same start configuration"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.11-1.png",
        "caption": "Figure 4.11 Pinion joined with shaft (a) by press-fitting (b) by key (Reproduced by permission of KISSsoft AG)",
        "texts": [
            " For a more uniform cross-section and better casting conditions, the common rib design is carried out pursuant to Figure 4.10, where all required dimension ratios are given. The shape of the gear body mainly depends on the operating conditions, service life, load, material and dimensions, size of the series and the technology available. When it is necessary for a pinion to be made from very high-quality and expensive material, it is worked separately and joined with the shaft either by press-fitted joints (Figure 4.11a), by spline joints or by key joints (Figure 4.11b). The design where the pinion is made in one piece with the shaft is recommended (Figure 4.12), because this design offers considerable advantages. It reduces the amount of machining (working the keyways on the gear hub and the shaft) and increases the shaft stiffness, as well as the stability of the wheel gear position. Steel gear wheels of a diameter less than 500mm are made of open-die or closed-die forgings (depending on the scale of production). Larger gears for larger series are either cast or split versions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001597_jae-2011-1343-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001597_jae-2011-1343-Figure1-1.png",
        "caption": "Fig. 1. Scheme of double-pendulum-type overhead crane.",
        "texts": [
            " Therefore, an individual that has less J is a better individual in GA. The real-valued GA that we implemented can be described algorithmically as follows: Step 1: Generate initial population, P (t) randomly. Step 2: If the number of generations <= maximum value, then go to Step3, otherwise Stop. Step 3: Evaluate each of the strings in P (t) using Eq. (18) according to the system simulation results. Step 4: Select and reproduction operation. Step 5: Crossover operation. Step 6: Mutation operation. In the section, the DPTOC system [8] illustrated in Fig. 1 is used to validate the system dynamics and the controller performance. Its model can be described as Eq. (2), where: q = [x1, x3, x5]T , \u03c4 = [F, 0, 0]T , M(q) = \u23a1 \u23a3 m + m1 + m2 (m1 + m2)l1 cos x3 m2l2 cos x5 (m1 + m2)l1 cos x3 (m1 + m2)l21 m2l1l2 cos(x3 \u2212 x5) m2l2 cos x5 m2l1l2 cos(x3 \u2212 x5) m2l 2 2 \u23a4 \u23a6 , C(q, q\u0307) = \u23a1 \u23a3 0 \u2212(m1 + m2)l1x\u03073 sin x3 \u2212m2l2x\u03075 sin x5 0 0 m2l1l2x\u03073 sin(x3 \u2212 x5) 0 \u2212m2l1l2x\u03073 sin(x3 \u2212 x5) 0 \u23a4 \u23a6 , G(q) = [0 (m1 + m2)gl1 sin x3 m2gl2 sin x5]T . In DPTOC, the system parameters that always change in different transport tasks are the payload mass m2 and the cable length l1 while those parameters like m1 and l2 may vary sometimes because the payload volume changes",
            " According to the theory of sliding mode control, the stability of the closed loop system consists of the stability in the sliding mode surface and the accessibility to the sliding mode surface. All the sliding mode surfaces s1 = 0, s2 = 0 and s3 = 0 are stable for \u03bb1, \u03bb2 and \u03bb3 are positive real numbers. The stability in the composite sliding mode surface is decided by the coupling factors k1, k2 and k3 in Eq. (16) and the definitions of the swing angles and the control input u (i.e. F ). Assume that positive angles are measured counterclockwise from the vertical line and positive control input u is towards the positive X-direction as in Fig. 1. When the position sliding mode function s1 > 0, a negative control input is required such that the sliding function s1 approaches the sliding mode surface s1 = 0. When the payload sliding mode function s3 > 0, a negative control input is required such that the sliding function s3 approaches the sliding mode surface s3 = 0. However, when the hook sliding mode function s2 > 0, a positive driving force is required such that the sliding function s2 approaches the sliding mode surface s2 = 0. Therefore, the role of the hook sliding mode function in control is contradictive with the others"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-Figure9-1.png",
        "caption": "Figure 9. Typical Simple Planetary GRA",
        "texts": [
            " Another option is to have two different gears (non-straight-cut) on a compound planet instead of a straight-cut plant gear. This arrangement has the same advantages and disadvantages of the previous arrangement except with a larger diameter and a shorter length. When a high-efficiency GRA is desired, a simple planetary gearbox is the better choice among the compound gear actuators. All the previous actuators have lower efficiency due to the differential ratio between two compound gear meshes. Cross section of a typical simple planetary GRA is shown in Figure 9, and the schematic is shown in Figure 10. The main disadvantage comes from the fact that it has more parts, which means lower reliability and higher cost to achieve the same gear ratio and torque capacity as the other arrangements. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 3 Copyright \u00a9 2011 by ASME Trailing Edge Rotary Actuators Flaps are hinged surfaces on the trailing edge of the wings. As flaps are extended with Fowler motion to increase the camber and the lift coefficient, this allows the aircraft to generate higher lift at slower speeds"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001531_s00707-013-0995-y-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001531_s00707-013-0995-y-Figure1-1.png",
        "caption": "Fig. 1 N-fold beam system (Here: N = 2, plane motion)",
        "texts": [
            " Introduction Figure 1 gives an idea of the systems under consideration. Here, vi (xi , t) represents the bending deflection in yi -direction w.r.t. the local (x\u2013z)i -frames, i = 1, 2. The \u201crigid body motions\u201d \u03b3i around the local zi -axes are assumed to be large. Subindex L denotes xi = Li with Li : length of i th beam, ( )\u2032 represents spatial derivative (w.r.t. x), and \u02d9( ) time derivative, respectively. Along with the mass distribution (\u2202m/\u2202x) and the inertia distribution (\u2202 J/\u2202x), one obtains the forearm bending equation (index 2): \u2202m2 \u2202x2 v\u03082 \u2212 \u2202 \u2202x2 [ \u2202 J2 \u2202x2 v\u0308\u2032 2 ] + \u22022 \u2202x2 2 [ (EI)2v \u2032\u2032 2 ] + \u2202m2 \u2202x2 { [ L1 cos \u03b32 \u2212 (L1v \u2032 1L \u2212 v1L ) sin \u03b32 + x2 ] \u03b3\u03081 + x2\u03b3\u03082 + [L1 sin \u03b32+(L1v \u2032 1L \u2212 v1L ) cos \u03b32\u2212v2 ] \u03b3\u0307 2 1 \u2212v2\u03b3\u0307 2 2 +[2v\u03071L sin \u03b32\u22122v2\u03b3\u03072 ] \u03b3\u03071 + g [ sin(\u03b31+\u03b32)+v\u2032 1L cos(\u03b31+\u03b32) ]+ (v\u03081L cos \u03b32 + x2v\u0308 \u2032\u2032 1L ) } \u2212 \u2202 \u2202x2 [ (\u03b3\u03081 + \u03b3\u03082 + v\u0308\u2032 1L ) \u2202 J2 \u2202x2 ] \u2212 \u2202 \u2202x2 \u23a1 \u23a3 L2\u222b x2 [ g cos(\u03b31+\u03b32)\u2212L1\u03b3\u03081 sin \u03b32+L1\u03b3\u0307 2 1 cos \u03b32+\u03be(\u03b3\u03071+\u03b3\u03072) 2] \u2202m2 \u2202\u03be d\u03be v\u2032 2 \u23a4 \u23a6 = 0",
            " The vector of describing velocities is known with Eq. (4) y\u0307 = ( vox voy \u03c9oz v\u0307 v\u0307\u2032 v\u0307\u2032\u2032 )T , (21) and the kinematic chain y\u0307i = Ti p y\u0307p + y\u0307i,rel reads \u239b \u239c\u239c\u239c\u239c\u239c\u239d vox voy \u03c9oz v\u0307(x, t) v\u0307\u2032(x, t) v\u0307\u2032\u2032(x, t) \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 i = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 Ai p \u23a1 \u23a31 0 \u2212vL 0 0 0 0 1 L 1 0 0 0 0 1 0 1 0 \u23a4 \u23a6 O \u2208 IR3,6 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 p \u239b \u239c\u239c\u239c\u239c\u239c\u239d vox voy \u03c9oz v\u0307L v\u0307\u2032 L v\u0307\u2032\u2032 L \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 p + \u239b \u239c\u239c\u239c\u239c\u239c\u239d vx,rel vy,rel \u03c9z,rel v\u0307(x, t) v\u0307\u2032(x, t) v\u0307\u2032\u2032(x, t) \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 i (22) where Ai p = \u239b \u239d cos \u03b3i p sin \u03b3i p 0 \u2212 sin \u03b3i p cos \u03b3i p 0 0 0 1 \u239e \u23a0 (23) (notice \u03b3i p = v\u2032 p L + \u03b3i see Fig. 1). Same as in Eq. (18), one obtains \u239b \u239c\u239c\u239d y\u03071 y\u03072 ... y\u0307N \u239e \u239f\u239f\u23a0 = \u23a1 \u23a2\u23a2\u23a3 O T21 O ... ... . . . TN1 TN2 \u00b7 \u00b7 \u00b7 O \u23a4 \u23a5\u23a5\u23a6 L \u239b \u239c\u239c\u239d y\u03071r y\u03072r ... y\u0307Nr \u239e \u239f\u239f\u23a0 L + \u239b \u239c\u239c\u239d y\u03071r y\u03072r ... y\u0307Nr \u239e \u239f\u239f\u23a0 , (24) index rel abbreviated r . However, the main difference here is that all those terms that are marked by L do not depend on the spatial variable x while the last term is a function of x (and t , of course). Inserting Eq. (24) into N\u2211 i=1 \u222b Li \u03b4yT i [dQ]c i = 0 (25) yields \u239b \u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239d \u222b \u03b4y1r\u222b \u03b4y2r\u222b \u03b4y3r ...\u222b \u03b4yNr \u239e \u239f\u239f\u239f\u239f\u239f\u239f\u239f\u23a0 T L \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 O TT 21 TT 31 \u00b7 \u00b7 \u00b7 TT N1 O TT 32 \u00b7 \u00b7 \u00b7 TT N2 O \u00b7 \u00b7 \u00b7 TT N3 ",
            " Acknowledgments Support of the Austrian Center of Competence in Mechatronics (ACCM) is gratefully acknowledged. Appendix: Elastic double pendulum, N = 2 A1. Equations of motion According to Eqs. (34) and (35), the equations of motion are derived from ( DT 1 \u25e6 [dQ/dx]c el,1 DT 2 \u25e6 [dQ/dx]c el,2 ) = ( 0 0 ) , \u23a1 \u23a3 [FT 1 ]rb [FT 1 TT 21]rb O [FT 2 ]rb \u23a4 \u23a6 L \u239b \u239c\u239d \u222b B1 [dQ]c rb,1\u222b B2 [dQ]c rb,2 \u239e \u239f\u23a0 = ( 0 0 ) . (A.1) The generalized distributed forces. Upper arm (i = 1) From Eq. (6) et seq. one obtains d Qc 1 = [My\u0308 + Gy\u0307 \u2212 Q]1 with y\u0307 according to Eq. (21) [where vox,1 = 0, voy,1 = 0, \u03c9oz,1 = \u03b3\u03071, see Fig. 1]: \u23a1 \u23a3 [dQ]c rb,1 [dQ]c el,1 \u23a4 \u23a6 = \u239b \u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239c\u239d \u2212(\u03b3\u03081v1 + x1\u03b3\u0307 2 1 + 2v\u03071\u03b3\u03071)dm1 \u2212 d fx (\u03b3\u03081x1 \u2212 v1\u03b3\u0307 2 1 + v\u03081)dm1 \u2212 d fy v\u03081x1dm1 + \u03b3\u03081dJ o 1 + v\u0308\u2032 1dJ1 \u2212 dMz + v1d fx \u2212 x1 d fy (\u03b3\u03081x1 \u2212 v1\u03b3\u0307 2 1 + v\u03081)dm1 \u2212 d fy (\u03b3\u03081 + v\u0308\u2032 1)dJ1 + [ \u222b L1 x1 ( d f (o) x d\u03be ) d\u03be ] v\u2032 1dx1 \u2212 dMz [(EI)1] v\u2032\u2032 1 dx1 \u239e \u239f\u239f\u239f\u239f\u239f\u239f\u239f\u239f\u239f\u239f\u239f\u23a0 (A.2) (second-order terms neglected). Comparing [dQ]c el,1 with the results from Table 1 (first row) shows that the accelerations are now completed with those arising from rigid body motion",
            " Exemplarily, we focus on [dQ]c el,2: From Eq. (6) et seq. (\u201clower part\u201d) one obtains [dQ]c el,2 = [dMT 12y\u0308rb + dM22y\u0308el + dG21y\u0307rb + dK22yel \u2212 dQ2 ] 2 (A.3) where y\u0307el,2 = \u239b \u239d v\u03072 v\u0307\u2032 2 v\u0307\u2032\u2032 2 \u239e \u23a0, y\u0307rb,2 = A21 \u239b \u239d \u2212v1L \u03b3\u03071 +L \u03b3\u03071 + v\u03071L \u03b3\u03071 + \u03b3\u03072 + v\u0307\u2032 1L \u239e \u23a0 (A.4) (see Eq. (22) along with vox,1 = 0, voy,1 = 0, \u03c9oz,1 = \u03b3\u03071 and [\u03c9z,rel]2 = \u03b3\u03072). For dMz = 0, the impressed force vector reads dQT 2,2 = (d fy, 0, 0). Considering gravitation, one has d fy = d fy,grav = \u2212dm2g sin \u03b32o, or, with \u03b32o = \u03b31 + v\u2032 1L + \u03b32 (see Fig. 1), d fy \u2212dm2g[sin(\u03b31 + \u03b32) + v\u2032 1L cos(\u03b31 + \u03b32)]. The zero-order reaction d f (o) x , which enters dK22, is obtained from \u2212eT 1 [dQ]c,o rb,2 (index o indicates that all deflections are zero, e1 \u2208 IR3: first unit vector), thus d f (o) x = \u2212e1 [ dM(o) 11 y\u0308(o) rb + dG(o) 11 y\u0307(o) rb \u2212 dQ(o) 1 ] 2 = \u2212dm2 [ L1\u03b3\u03081 sin \u03b32\u2212L1\u03b3\u0307 2 1 cos \u03b32\u2212x(\u03b3\u03071+\u03b3\u03072) 2 \u2212 d f (o) x,grav ] where d f (o) x,grav = dm2g cos(\u03b31 + \u03b32). Setting dm2 = (\u2202m2/\u2202x2)dx2 yields [dQ]c el,2 = [(dQ/dx)dx]c el,2 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 \u2202m2 \u2202x2 v\u03082 { + [ L1 cos \u03b32 \u2212 (L1v \u2032 1L \u2212 v1L ) sin \u03b32 + x2 ] \u03b3\u03081 + x2\u03b3\u03082 + [ L1 sin \u03b32 + (L1v \u2032 1L \u2212 v1L ) cos \u03b32 \u2212 v2 ] \u03b3\u0307 2 1 \u2212 v2\u03b3\u0307 2 2 + [2v\u03071L sin \u03b32 \u2212 2v2\u03b3\u03072 ] \u03b3\u03071 + g [ sin(\u03b31 + \u03b32) + v\u2032 1L cos(\u03b31 + \u03b32) ] + (v\u03081L cos \u03b32 + x2v\u0308 \u2032 1L ) } \u2202m2 \u2202x2 (\u03b3\u03081 + \u03b3\u03082 + v\u0308\u2032 1L + v\u0308\u2032 2) \u2202 J2 \u2202x2 + L2\u222b x2 [ g cos(\u03b31+\u03b32)\u2212L1\u03b3\u03081 sin \u03b32+L1\u03b3\u0307 2 1 cos \u03b32+\u03be(\u03b3\u03071+\u03b3\u03072) 2 ] \u2202m2 \u2202\u03be d\u03be v\u2032 2 (EI)2v \u2032\u2032 2 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 dx2 = 0"
        ],
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    {
        "image_filename": "designv11_100_0003563_amr.311-313.360-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003563_amr.311-313.360-Figure1-1.png",
        "caption": "Figure 1 Thrust bearing or seal moving coordinate system",
        "texts": [
            " In this paper, under the hypothesis of small perturbation approximation, the dual number was used in the dynamic property analysis of gas-lubricated thrust bearings or non-contacting face seals, and the perturbed Reynolds\u2019 equations was obtained, which is independent of its whirl frequency, that is independent of the system motion equations, and then through easily solving the perturbed pressure m p m p \u2202 \u2202 \u2202 \u2202 , \uff08m=z,\u03b1,\u03b2\uff09equations of the bearings\u2019 or seals\u2019 dynamic gas film to calculate the stiffness and damping coefficients which determine the dynamic operation stability of the system. 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: 152.14.136.77, NCSU North Carolina State University, Raleigh, USA-23/04/15,11:18:29) Approximated Perturbed Reynolds\u2019 Equations of gas film In dimensionless cylindrical ACS (R,\u03b8,T), the compressible Reynolds equation is [6] : (see Figure1) ] )( 2 )( [2)()( 2 3 2 3 2 T PH \u03b8 PH \u039b R P RH RR\u03b8 P H \u03b8R \u2202 \u2202 + \u2202 \u2202 = \u2202 \u2202 \u2202 \u2202 + \u2202 \u2202 \u2202 \u2202 (1) Where , ap p P = , ch h H = , ir r R = \u03c9t=T \u039b\u2500\u2500 Harrison Number 2 c 26 hp \u00b5\u03c9r \u039b a i= \u03bc\u2500\u2500 Dynamic viscosity pa\u2500\u2500 Standard State Pressure \u03c9\uff0cri\uff0chc\u2500\u2500 running angular velocity (frequency) , minimum radius, the maximum steady-state film thickness of the dynamic wedge. Now, expressed in the form of dual number, the perturbed displacement of dimensionless axial Z and dimensionless angular A, B under the critical state are: )(exp)(1()( Z0Z0 T\u03b3ZT\u03b3ZTZ \u2208=\u2208+= (2a) )(}exp)(1}(} )( )( N 0 0 N 0 0 T\u03b3 B A T\u03b3 B A TB TA \u2208=\u2208+= (2b) Where , ch z Z = \u03b2 \u03b1 h r B A i {} c = Z0\uff0cA0\uff0cB0\u2500\u2500 initial value of axial, angular perturbation \u03b3Z,\u03b3N\u2500\u2500 whirl frequency ratio of axial, angular perturbation \u03c9 \u03bd \u03b3 \u03c9 \u03bd \u03b3 n N, == z Z \uff08the correlation of angular motion\uff09 \u03bdz\uff0c\u03bdn\u2500\u2500 critical self-excited whirl frequency of bearing or seal The dual number parameter\u03b3T express the feature of the dual number Z and A, B, namely the nature of the perturbation displacement, T\u2208[0\uff0c\u221e]"
        ],
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    },
    {
        "image_filename": "designv11_100_0003710_icca.2011.6138010-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003710_icca.2011.6138010-Figure6-1.png",
        "caption": "Fig. 6. Model of the hovercraft. Body fixed XBYB and earth fixed coordinate frames XY.",
        "texts": [
            " With this purpose, the control problem is tackled as a multivariable nonlinear control design using QFT technique. The approach to non-linear QFT synthesis follows the ideas described in [34], where a local linearization of the nonlinear plant about closed-loop acceptable outputs is proposed. A. Model of the hovercraft The nonlinear model for the underactuated hovercraft was obtained from the ship model in [35]. The general kinematic and dynamic equations of motion of the hovercraft can be developed using a global coordinate frame {XY} and a body fixed coordinate frame {XBYB} that are depicted in Fig. 6. Considering that the state vector is the non-linear state equations are [36]: rr J T J r vr m urv ur m F m vru r vuy vux r l lx \u00b7\u00b71\u00b71 \u00b7\u00b71\u00b7 \u00b7\u00b71\u00b71\u00b7 \u00b7cos\u00b7sin \u00b7sin\u00b7cos (19) where x, y, denote the position and the orientation of the hovercraft in the earth-fixed frame XY; u, v are respectively the surge and sway velocities in the body-fixed frame XBYB, and r is the yaw rate. The system has two control inputs: Fx = (Fs + Fp) is the control force in surge, and T = l\u00b7(Fs-Fp) is the control torque in yaw",
            " The hovercraft nominal parameters have been computed experimentally in a real system: mass m = 0.894 Kg, moment of inertia J = 0.0125 Kg\u00b7m2, moment arm l = 0.0485 m, friction coefficients rl = 0.10 Kg/s, and rr = 0.05 Kgm2s, and F [0.342, -0.121] N. The control objective is to achieve the tracking control. The two outputs are: the tangential velocity V, defined as 22 vuV , and the derivative of the course angle , defined as the angle that the tangent of the trajectory in the XBYB plane makes with the inertial X-axis (see Fig. 6), that is, )/arctan( xy , also defined from the attitude angle and offset angle in yaw as . a c a c f p p n n t n l g w E s s c F d i c f s a c t w i t For the stat nd velocity ircumference peak oversh ontains plant riction coeff articular QFT B. Nonlinea 1) Step 1. roblem of a onlinear sys onlinear QFT [34]. This tec heorem [15] onlinear plan inear family iving as a re ill be solved 2) Step 2. L LF is compu ystem (Fig. 7 {P} obtained ystems. Als onsidered. A rom fixed po esign of two nput, singleonsists of the ii for the plan uch that sat chieve the tra Fig"
        ],
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    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.12-1.png",
        "caption": "Fig. 2.12 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRRP (a) and 4RRRPR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R||R||R||P (a) and R\\R||R||P||R (b)",
        "texts": [
            "1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig. 2.9a) R||R||P||R\\R (Fig. 2.1n) Idem No. 13 16. 4RRRPR (Fig. 2.9b) R||R||R||P\\R (Fig. 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No. 13 18. 4RRPRR (Fig. 2.10b) R||R||P||R\\R (Fig. 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2.1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2"
        ],
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    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.48-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.48-1.png",
        "caption": "Fig. 2.48 4PaPaPR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\Pa\\\\P\\\\R",
        "texts": [
            "42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig. 2.49) R\\Pa||Pa\\kR (Fig. 2.22j) The first and the last revolute joints of the four limbs are parallel 38. 4RPaPaR (Fig. 2.50) R\\Pa||Pa\\kR (Fig. 2.22k) Idem No. 37 39. 4RPaRPa (Fig. 2.51) R\\Pa\\kR\\Pa (Fig. 2.22l) Idem No. 15 40. 4PaRPPa (Fig. 2.52) Pa\\R\\P||Pa (Fig. 2.22m) Idem No. 34 41. 4PaPaRP (Fig. 2.53) Pa||Pa\\R\\P (Fig. 2.22n) The axes of the revolute joints connecting links 7 and 8 of the four limbs are parallel 42. 4PaPaPR (Fig. 2.54) Pa||Pa||P\\R (Fig"
        ],
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    },
    {
        "image_filename": "designv11_100_0003651_s0025654413030023-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003651_s0025654413030023-Figure1-1.png",
        "caption": "Fig. 1.",
        "texts": [
            " Kinematic models also do not allow one to study the forces acting on WS, for example, in motion along a slope, when towing loads, in the case of significant variations in the road friction coefficients, and under conditions of strong crosswind. In what follows, dynamic WS models are constructed with the inertial properties of the system taken into account in explicit form [18\u201323]. The goal is to study the influence of external forces and the WS mass distribution on the dynamic properties of the system and construct a WS control law on this basis. The system presented in Fig. 1, which is a model of a wheeled vehicle (like a wheeled tractor or a car), is studied. This WS contains the body, the driving rear axle, and the steering front axle. The inertial properties of the system are taken into account in the following simple approximation. The WS is considered as a solid of mass m and with the moment of inertia J with respect to the center of mass of the system at the point (xc, yc). To this end, it is assumed that the masses of the wheels and the front axle are sufficiently small and the velocities of motions of these masses are not large. The state of the WS body is characterized by the angle a and the position of the point p between the rear wheels; here (x, y) are coordinates of the point p (Fig. 1). The rear wheels are leading; i.e., a force w (a control) is applied to the rear axle. It is also assumed that the mechanical system is subjected to the constraints x\u0307 sin a \u2212 y\u0307 cos a = 0, x\u0307 sin(a + b) \u2212 y\u0307 cos(a + b) \u2212 La\u0307 cos b = 0. (1.1) The first relation in (1.1) reflects the assumption that the WS rear wheels do not slip in the direction along their axes, and the second relation reflects a similar assumption about the front wheels. The objective of *e-mail: v.matyuhin@mail.ru 243 the trajectory control problem is make the point p with coordinates (x, y) move along a given plane smooth curve S"
        ],
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        "image_filename": "designv11_100_0002783_1.3543587-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002783_1.3543587-Figure4-1.png",
        "caption": "Fig. 4. Single-spring magnetically driven oscillator arrangement.",
        "texts": [
            " Zein Nakhoda is a graduate of Lake Highlands High School in Dallas, TX (Richardson ISD) and currently attends Swarthmore College in Pennsylvania. Swarthmore College, Swarthmore, PA 19081; wnakhod1@ swarthmore.edu Ken Taylor teaches AP Physics B and C at Lake Highlands High School in Dallas, TX (Richardson ISD). Lake Highlands High School (Richardson ISD), 9449 Church Road, Dallas, TX 75238; ken.taylor@risd.org 750 Interface and Power Amplifier II to serve as an electromagnet acting on the magnets7 (see Fig. 4). The differential equation describing the motion of the system is the same as Eq. (7) with the exception of the inclusion of the driving force F(t). Hence, ( ) ( )2 2 m s m ( ).I mR kR mgR F t\u03b8 \u03b8+ + = + (11) Since the field acting on the magnet is not uniform and the magnet is dipolar, the actual force function between the magnet and electromagnet is not known. Hence, F(t) cannot be specified for Eq. (11). A specific waveform can, of course, be selected at the interface in order to give the coil current (and magnetic field) a specific form"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002841_05017.0001ecst-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002841_05017.0001ecst-Figure1-1.png",
        "caption": "Figure 1. Illustration of the three-electrode electrochemical cell. W.E., C.E., R.E. are the abbreviations of working, counter, and reference electrodes, respectively.",
        "texts": [
            " The electrochemical glass cell and the Pt counter electrode were rinsed with acetone, ethanol, and water, then dried and soaked in 50/50 H2SO4/HNO3 solution (Wako Pure Chemical Industries, Ltd.) at least overnight. The ITO nanoparticle 3 ) unless CC License in place (see abstract).\u00a0 ecsdl.org/site/terms_use address. Redistribution subject to ECS terms of use (see 128.218.248.200Downloaded on 2015-03-09 to IP electrode was encircled with an O-ring of 12 mm inner diameter, and then the electrochemical glass cell was set on the O-ring as shown in Fig. 1. After the glass cell was filled with the electrolyte solution, the counter and reference electrodes were dipped in the electrolyte. The current was collected by cupper adhesive tape sticking on the ITOfilm surface in the outer side of the O-ring. Cyclic voltammetry (CV) was carried out in the PBS solutions containing and not containing Hb and H2O2 at the ITO nanoparticle electrodes at room temperature. Evaluation of biofuel cells utilizing enzymatic activity of Hb at ITO nanoparticle electrodes The single cells assembled in this study were composed of homemade accessories as shown in Fig",
            " The resultant semi-MEA was set on the single cell anode as shown in Fig. 2. H2 gas can be provided to the backside of the Pt-dispersed carbon paper similarly to polymer electrolyte fuel cells (35,36). The center hole for the Teflon block facing the Pt-dispersed carbon paper was 10 mm in diameter. The ITO nanoparticles were deposited on the carbon paper, as is the case with the ITO-coated glass plate. The ITO nanoparticle electrode on carbon paper was faced with the Nafion membrane of semiMEA as shown in Fig. 1 and used as the cathode. The 10-mm-diameter center hole of 4 ) unless CC License in place (see abstract).\u00a0 ecsdl.org/site/terms_use address. Redistribution subject to ECS terms of use (see 128.218.248.200Downloaded on 2015-03-09 to IP Teflon block at the cathode side was filled with 5 \u00b5M Hb/1.8 mM H2O2 PBS solution. The single cell performances were evaluated by a slow current sweep technique on a Solartron SI1287 potentiostat/galvanostat. Characterization of prepared ITO nanoparticles Figure 3 shows the XRD patterns of the ITO product prepared here and from the ICDD card (#06-0416) for In2O3 with a cubic crystal system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002604_amr.482-484.2017-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002604_amr.482-484.2017-Figure5-1.png",
        "caption": "Fig. 5, Distribution of equivalent stresses in the roller and racer in the third statement of the task",
        "texts": [
            " In the third formulation contact pressures are precisely enough calculated according the Hertz theory along a larger part of the roller ends, but near faces (Fig. 3, c) you can observe sharp pressure which distribution takes the form of \"a dog bone\" [5]. Stressed state. Let`s consider pressure distribution inside the bearing ring (Fig. 4) expressed in dimensionless coordinates. Pressure distribution graphs in Fig 4 (at the left) confirm correctness of the Hertz-Belyaev problem solution about pressure distribution along the line contact [3,6,7] but an interesting effect at the roller edge is seen (Fig. 5, to the right): pressures max\u041aY \u041aP\u03c3 and max\u041aX \u041aP\u03c3 have practically the same values along the whole length of the ring. The curve of equivalent pressures max\u041aY \u041aP\u03c3 has sharper bend than in the roller centre max33 PREPR P\u03c3 . Concentration of equivalent pressures along the major part of the roller length is at a depth of 0.6, this value falling up to 0.3 at the edge. Pressure isolines for the third task formulation are shown in Fig. 4 where you can see increase of equivalent pressures at the roller edge"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002474_amr.614-615.1226-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002474_amr.614-615.1226-Figure3-1.png",
        "caption": "Fig. 3 FEA mesh of the novel HEBCA (a) Rotor. (b) Stator. (c) Excitation bracket.",
        "texts": [
            " Compared with the ordinary electric excitation claw-pole alternator, this hybrid excitation brushless claw-pole alternator reduces the excitation loss, increases the alternator\u2019s power density, and improves the alternator\u2019s low speed power characteristic. Furthermore, its excitation winding is stationary, thus the brush and slip ring can be eliminated. So it has good reliability. The magnetic field regulation for HEBCA In order to get the accurate magnetic field distribution and evaluate the field regulating performances of the novel hybrid excitation brushless claw-pole alternator, 3-D finite element analysis is performed on a prototype alternator. Fig. 2 shows the FEA model of the HEBCA, and Fig. 3 is the mesh of the HEBCA. Fig.1 Structure of the novel HEBCA Fig. 2 FEA model of the novel HEBCA (a) (b) Based on the model above, the no-load magnetic field distributions for different excitation currents are calculated, and the distributions of air-gap flux density are shown in Fig. 4. It can be seen that air-gap flux density is close to zero when the excitation current is 0, and that the air-gap flux density increase with the excitation current. Fig. 5 shows the average no-load air-gap flux density of HEBCA and EECA for different excitation currents"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.19-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.19-1.png",
        "caption": "FIGURE 1.19. The di erence between aluminum, magnesium, and steel rims in regaining road contact after a jump.",
        "texts": [
            " Aluminum is very good for its weight, thermal conductivity, corrosion resistance, easy casting, low temperature, easy machine processing, and recycling. Magnesium is about 30% lighter than aluminum, and is excellent for size stability and impact resistance. However, magnesium is more expensive and it is used mainly for luxury or racing cars. The corrosion resistance of magnesium is not as good as aluminum. Titanium is much stronger than aluminum and magnesium with excellent corrosion resistance. However, titanium is expensive and hard to be machine processed. The di erence between aluminum, magnesium, and steel rims is illustrated in Figure 1.19. Light weight wheels regain contact with the ground quicker than heavier wheels. Example 39 Spare tire. Road vehicles typically carry a spare tire, which is already mounted on a rim ready to use in the event of at tire. Since 1980, some cars have been equipped with spare tires that are smaller than normal size. These spare tires are called doughnuts or space-saver spare tires. Although the doughnut spare tire is not very useful or popular, it can help to save a little space, weight, cost, and gas mileage"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003224_amr.565.171-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003224_amr.565.171-Figure6-1.png",
        "caption": "Fig. 6 Temperature distributions of oil film at different oil supply pressures",
        "texts": [
            " When other factors remain unchanged, the change of oil supply pressure will lead a change about bearing flow. With the rising of oil supply pressure, the flow of bearing will increase, so more heat will be taken away, and the temperature of bearings will decrease. Under the condition, keeping eccentricity ratios 0.5\u03b5 = , rotational speed 10000 r/minv = and supply oil temperature o25 CT = unchanged, the temperature field distributions of oil film in different oil supply pressures is obtained by simulation analysis, which are shown as Fig. 6, and the law of temperature change of oil film is shown as Fig. 7. With the rising of oil supply pressure, the temperature of oil film declines continually, low-temperature area gradually expands outwards from oil inlet, and high-temperature area shrinks along the reverse direction. Fig. 7 shows that, the relationship of oil film temperature with oil supply pressure is inverse ratio. With the rising of oil supply pressure, both maximum temperature and average temperature of oil film decline, and the value of average temperature is infinitely close to the temperature of supplied oil"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.8-1.png",
        "caption": "Figure 12.8 Intersection of the momentum and energy ellipsoids.",
        "texts": [
            "162) Therefore, the dynamics of the moment-free motion of a rigid body requires that the corresponding angular velocity \u03c9(t) satisfy both Equations (12.161) and (12.162) and therefore lie on the intersection of the momentum and energy ellipsoids. For a better visualization, let us define the ellipsoids in the (Lx, Ly, Lz) coordinate system as L2 x + L2 y + L2 z = L2 (12.163) L2 x 2I1K + L2 y 2I2K + L2 z 2I3K = 1 (12.164) Equation (12.163) is a sphere and Equation (12.164) defines an ellipsoid with \u221a 2IiK as semiaxes. To have a meaningful motion, these two shapes must intersect. The intersection forms a trajectory, such as shown in Figure 12.8. We conclude that for a given angular momentum there are maximum and minimum limit values for possible kinetic energy. Assuming I1 >I3 >I3 (12.165) the limits of possible kinetic energy are Kmin = L2 2I1 (12.166) Kmax = L2 2I3 (12.167) and the corresponding motions are turning about the axes I1 and I3, respectively. Example 727 Alternative Derivation of Euler Equation of Motion Consider a rigid body with a fixed point as shown in Figure 12.9. A small mass dm at Grdm is under a small force df. Let us show the moment of the small force df by dm: Gdm = Grdm \u00d7 Gdf = Grdm \u00d7 Gv\u0307dm dm (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002656_wcse.2013.52-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002656_wcse.2013.52-Figure2-1.png",
        "caption": "Fig. 2. relative airflow of rotor.",
        "texts": [
            " If QUAV-X4 is hovering, when augmented or abated rotating speeds of four rotors simultaneously, QUAV-X4 flies up and down vertically, and it will be rolling, pitching, or yawing when varying rotating speeds of one or more rotors, then because of coupling, it will be flying linearly in the space. The roll angle can be controlled indirectly, the same as the pitch angle . The yaw angle can be controlled by relatively changing the rotating speeds of two group rotors. II. AERODYNAMIC MODEL For designing aerodynamic model of QUAV-X4, the key is to obtain aerodynamic coefficients, and those coefficients can be derived by blade element theorem. Rotor structure shaft system os-xsyszs is drawn as Fig. 2. The speed of relative airflow is V0, s is attack angle of rotor, is induced speed of rotor, which is derived by momentum theorem [4]. By analysis of relative airflow of rotor, when QUAV-X4 is in flight, rotor creates thrust T, twist torque Q, drag D, and heeling moment H. These forces and torque are proportional to the square of rotating speed of rotor, so, T=CT A( R)2. (1) D=CD A( R)2. (2) Q=CQ A( R)2 . (3) H=CH A( R)2 . (4) CT, CD, CQ and CH are aerodynamic coefficients, means angular speed of rotor, R is rotating radius of rotor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001736_978-3-319-01851-5_12-Figure12.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001736_978-3-319-01851-5_12-Figure12.7-1.png",
        "caption": "Fig. 12.7 Rolling robot with ODWs in a -array",
        "texts": [
            " On the other hand, the determination of the motion produced by a given history of joint torques requires (a) the calculation of I, which can be achieved symbolically; (b) the inversion of I, which can be done symbolically because this is a 2 2 matrix; (c) the calculation of the Coriolis and centrifugal terms, as well as the dissipative forces; and (d) the integration of the initial-value problem resulting once initial values to and P a have been assigned. We now consider a three-dof robot with three actuated wheels of the Mekanum type, as shown in Fig. 10.19, with the configuration of Fig. 12.7, which will be termed, henceforth, the -array. This system is illustrated in Fig. 12.8. Below we will adopt the notation of Sect. 10.5.2, with \u02db D =2 and n D 3. We now recall that the twist of the platform was represented in planar form as t0 ! Pc (12.55) where ! is the scalar angular velocity of the platform and Pc is the two-dimensional position vector of its mass center, which will be assumed to coincide with the -array centroid of the set of points fCi g31. Moreover, the three wheels are actuated, and hence, the three-dimensional vector of actuated joint rates is defined as P a 2 4 P 1P 2P 3 3 5 (12.56) The relation between P a and t0 was derived in general in Sect. 10.5.2. As pertaining to the robot of Fig. 12.7, we have J P a D Kt0 (12.57a) with the two 3 3 Jacobians J and K defined as J a1; K 2 4r fT1 r fT2 r fT3 3 5 (12.57b) where, it is recalled, a is the height of the axis of the wheel hub and r is the horizontal distance of the points of contact with the ground to the mass center C of the platform, as indicated in Fig. 12.7a. Moreover, vectors f ei g31 and f fi g31, defined in Sect. 10.5.2, are displayed in Fig. 12.7. Below we derive expressions for ! and Pc, from Eq. (12.57a), in terms of the joint rates. To this end, we expand these three equations, thus obtaining r! C fT1 Pc D a P 1 (12.58a) r! C fT2 Pc D a P 2 (12.58b) r! C fT3 Pc D a P 3 (12.58c) Upon adding corresponding sides of the three foregoing equations, we obtain 3r! C PcT 3X 1 fi D a 3X 1 P i (12.59) But from Fig. 12.7b, it is apparent that e1 C e2 C e3 D 0 (12.60a) f1 C f2 C f3 D 0 (12.60b) Likewise, e1 D p 3 3 .f3 f2/; e2 D p 3 3 .f1 f3/; e3 D p 3 3 .f2 f1/ (12.60c) f1 D p 3 3 .e2 e3/; f2 D p 3 3 .e3 e1/; f3 D p 3 3 .e1 e2/ (12.60d) and hence, the above equation for ! and Pc leads to ! D a 3r 3X 1 P i (12.61) Now we derive an expression for Pc in terms of the actuated joint rates. We do this by subtracting, sidewise, Eq. (12.58b) from Eq. (12.58a) and Eq. (12.58c) from Eq. (12.58b), thus obtaining a system of two linear equations in two unknowns, the two components of the two-dimensional vector Pc, namely, APc D b with matrix A and vector b defined as A .f1 f2/T .f2 f3/T p 3 eT3 eT1 ; b a P 1 P 2P 2 P 3 where we have used relations (12.60c). Since A is a 2 2 matrix, its inverse can be readily found with the aid of Facts 5.7.3 and 5.7.4, which yield Pc D 2 3 a Ee1 Ee3 P 1 P 2P 2 P 3 Now, from Fig. 12.7b, Ee1 D f1; Ee3 D f3 and hence, Pc reduces to Pc D 2 3 a\u0152. P 2 P 1/f1 C . P 2 P 3/f3 2 3 a\u0152 P 2.f1 C f3/ P 1f1 P 3f3 But by virtue of Eq. (12.60b), f1 C f3 D f2 the above expression for Pc thus becoming Pc D 2a 3 3X 1 P i fi (12.62) Thus, ! is proportional to the mean value of f P i g31, while Pc is proportional to the mean value of f P i fi g31. In deriving the mathematical model of the robot at hand, we will resort to the natural orthogonal complement, and therefore, we will require expressions for the twists of all bodies involved in terms of the actuated wheel rates",
            " Henceforth, we will regard the angular velocity of the platform and the velocity of its mass center as three-dimensional vectors. Therefore, t4 T4 P a; T4 k k k 2rf1 2rf2 2rf3 (12.63) with defined, in turn, as the ratio a 3r (12.64) Now, the wheel angular velocities are given simply as !i D P iei C !k D P iei 3X 1 P i ! k (12.65) or !1 D .e1 k/ P 1 P 2k P 3k (12.66a) !2 D P 1k C .e2 k/ P 2 P 3k (12.66b) !3 D P 1k P 2k C .e3 k/ P 3 (12.66c) Similar expressions are derived for vectors Pci . To this end, we resort to the geometry of Fig. 12.7, from which we derive the relations Pci D Pc C !rfi D 2 r 3X 1 P j fj ! r 3X 1 P j ! fi and hence, Pc1 D r\u0152.3 P 1 C P 2 C P 3/f1 C 2. P 2f2 C P 3f3/ (12.67a) Pc2 D r\u01522 P 1f1 C . P 1 C 3 P 2 C P 3/f2 C 2 P 3f3 (12.67b) Pc3 D r\u01522. P 1f1 C P 2f2/C . P 1 C P 2 C 3 P 3/f3 (12.67c) From the foregoing relations, and those for the angular velocities of the wheels, Eqs. (12.66a\u2013c), we can now write the twists of the wheels in the form ti D Ti P a; i D 1; 2; 3 (12.68) where T1 e1 k k k 3 rf1 r.f1 C 2f2/ r"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003699_s0036029511040239-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003699_s0036029511040239-Figure2-1.png",
        "caption": "Fig. 2. Schematic diagram of the residual stresses operat ing in a wire core and the glass shell.",
        "texts": [
            " 4 ESTIMATION OF THE INTERNAL STRESSES IN AMORPHOUS \u2026 MICROWIRES 377 During indentation, a complex state of stress close to uniform compression is created near the contact zone. The deformation penetrating deep into the material has both elastic and plastic components, which makes it possible to obtain information on both the hardness and Young\u2019s modulus [7]. Using the solution to the Lam\u00e9 problem of axisymmetric deformation of a tube, we determined the radial, tangential, and axial stresses using the following formulas (Fig. 2) [8]: for a glass shell, we used for a metallic core, we used Here, D is the outer diameter of the shell; d is the core diameter; \u03b4 = (\u03b10 \u2013 \u03b1c)(T \u2013 T0); \u03d5 = 1 + D2/d2; \u03c9 = E0/Ec(1 \u2013 D2/d2); E0 and Ec are Young\u2019s modulus of the glass and metallic core, respectively; \u03b10 and \u03b1c are the TECs of the glass and metallic core, respectively; T0 is the environment temperature; T is the glass soft ening temperature; and \u03bd is the Poisson ratio. To analyze the effect of the metallic core size and glass shell thickness t on the internal stresses, we chose the following microwires: in the former case, microw ires had approximately the same glass shell thickness; in the latter case, microwires had the same metallic core diameter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003784_s0036029512050060-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003784_s0036029512050060-Figure1-1.png",
        "caption": "Fig. 1. Schematic diagram of deformation during the compression of two metallic sheets 1 and 2 with tilted grooves 3\u20136 made on the surface of one of them (see text).",
        "texts": [
            " This factor depends on the ratio of their atomic radii, their wetting characteristics, and the temperature. The purpose of this work is to ensure a high bimetal strength when the requirements imposed on the clean ing and heating of contact surfaces are softened. In the new technological process [7], two plates are subjected to joint plastic deformation. Grooves, which are tilted to a contact surface in various directions (including opposite directions), are formed on the contact surface of at least one plate. As an illustration, Fig. 1 shows two metallic sheets 1 and 2 (dashed line DOI: 10.1134/S0036029512050060 RUSSIAN METALLURGY (METALLY) Vol. 2012 No. 5 NEW METHOD FOR THE PRODUCTION OF COMPOSITE MATERIALS 363 depicts the shape of sheet 1 before the beginning of deformation). In sheet 2, grooves 3 and 4 were prelim inarily made on the left of axis y and grooves 5 and 6 were made on the right of axis y; they have different slopes to the contact surface. Since grooves 3\u20136 are tilted to the contact surface (plane A) and axis y, the metal of sheet 1 fills them partly after joint plastic deformation during rolling or compression, and permanent joint of sheets 1 and 2 is achieved [7, 8]",
            " Due to this joint, the sheets cannot move along axes x and y, since the grooves are not par allel to each other and a strong contact appears after partial filling with the metal of sheet 1 (which has a lower strength). As a result, part of the sheet 1 metal forms keys that ensure a strong joint. It is convenient to use an even number of grooves having slopes to axis y that have the same modulus and opposite directions. In [7], we recommended to use angles of groove inclination of 45\u00b0\u201370\u00b0 to contact sur face A, i.e., angles \u03b1 = 20\u00b0\u201345\u00b0 to the normal to the surface. In the case shown in Fig. 1, we have \u03b1 = \u00b130\u00b0 for grooves 3 and 6 and \u03b1 = \u00b120\u00b0 for grooves 4 and 5. For every pair of grooves, the slopes have the same absolute value and opposite directions. If grooves also have the same absolute value and opposite angles of inclination to the third axis (z), which is normal to axes x and y and the plane of Fig. 1, these grooves prevent displace ment along this axis. Although grooves 3\u20136 can be machined, it is better to use extrusion, i.e., metal forming. The sheets do not undergo displacement due to both the seizure of the joint and the strength of projec tions, which partly fill the grooves, that can exclude hot rolling and rolling in vacuum or an inert gas in some cases. When it is necessary to join two sheets one which is thin (which is often necessary to produce molds: it is necessary to join a copper sheet 4\u20135 mm thick to a steel sheet 15\u201325 mm thick), one can produce a sheet with increased thickness (projection) regions (see Fig",
            "95 of the corresponding groove width in sheet 2 in order to facilitate filling of the grooves. The position of sheet 1 after deformation is shown by the dashed line in Fig. 2. In this case, the deformation of sheets 1 and 2 in rolling (or upsetting) can be decreased and deep grooves can be filled. Although projections can be formed by milling or planning, it is better to use metal forming, e.g., rolling in grooved rolls. To estimate the allowable shear force when, e.g., sheet 1 moves with respect to sheet 2 along the x axis (Fig. 1), we should take into account that shear can occur along plane A and the sheet 1 metal with a lower strength can fail. In this case, the fracture force is where \u03c4u1 is the ultimate strength during shear of the sheet 1 metal, n is the number of grooves, a is the groove width along the x axis, and L is the bimetal sheet length perpendicular to the drawing plane in Fig. 1. T1 \u03c4u1naL,= 364 RUSSIAN METALLURGY (METALLY) Vol. 2012 No. 5 BROVMAN Moreover, fracture can also occur in the sheet 2 metal in shearing along plane B in Fig. 1. In this case, the shear force is where \u03c4u2 is the ultimate strength during shear of the sheet 2 metal and b is the total width. The type of fracture depends on which of quantities T1 and T2 is lower. Of course, it is better if the equal strength condi tion T1 = T2 is met (otherwise, the strength of one of the joined metals is incompletely used). As a result, we have T2 \u03c4u2 b na\u2013( )L,= \u03c4u1naL \u03c4u2 b na\u2013( )L,= na \u03c4u 1b \u03c4u1 \u03c4u 2+ .= We now take into account that, for many metals, ulti mate shear strength is related to ultimate tensile strength by the relationship and obtain (1) The limiting force is (2) under these conditions, T1max = T2max and the equal strength condition is met",
            " The disadvantage of a number of well known methods for producing composite materials is the fact that, if even a small crack appears in a contact layer or near it, it can easily grow and causes fracture under alternating stresses. Such a small crack can be repre sented by an air bubble, a nonmetallic inclusion, a scale piece, and so on. Defects are dangerous even at a very low content. However, in the method of joint plas tic deformation with filled tilted grooves, a crack having appeared in the contact zone between grooves 3 and 4 (Fig. 1) can propagate only to these grooves (similarly, cracks between grooves 4 and 5 or 5 and 6 can grow only in the zone between neighboring grooves). This is the advantage of this method. Of course, in some cases this method cannot replace other methods, such as welding or joint rolling in vacuum [7]. Nevertheless, it can be sometimes used for joining metals that can hardly be joined by welding, and rolling in vacuum requires much energy. We assume the simplest kinematically admissible velocity field for the compression deformation of a plate of thickness h and width L under plane deforma tion conditions in the presence of two tilted grooves"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003659_amr.228-229.106-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003659_amr.228-229.106-Figure7-1.png",
        "caption": "Fig. 7 Contact lines in the plane ABCD Fig. 8 A pair of cylindrical spiral involute gears According to the gearing principle of involute gears, the contact line between helical gear 1 and plane ABCD moves uniformly from P0Q0 to PtQt. Simultaneously the contact line between helical gear 2 and plane ABCD moves uniformly from R0S0 to RtSt (See Fig. 7). For gear 1-gear 2 set, contact line PQ and RS always intersects in the plane ABCD. The intersection of them moves uniformly from B to D. Hence, line BD is the trajectory of contact for the right profile of gear 1 and the left profile of gear 2. Actually, a spiral involute on the right profile of gear 1 is always contact tangentially with the spiral involute on the left profile of gear 2. In light of gearing principle for involute rack-pinion, once gear 1 rotates uniformly, plane ABCD, the virtual common-rack-cutter surface will move along the tangent to pitch circle of the gear 1, and simultaneously gear 2 must rotate uniformly driven by plane ABCD. On the other side of the tooth is the virtual common surface EFGH with different hand from plane ABCD. Therefore hands of right-left profiles for gear 1 or gear 2 are different. So the slot width or tooth thickness varies along the axis of the gear, which looks like a",
        "texts": [
            " Axis x coincides with axis x , and axis z is perpendicular to axis z . Suppose \u03a3 revolves about axis z at angular velocity 1\u03c9 and \u03a3\u2032 revolves about axis z at angular velocity 2\u03c9 . Unit normal vector of involute helicoid \u03a3 is: kCjBiAN 0000 ++= (8) Where ( ) ( ) ( ) ( ) 22 0 0 22 00 2 0 0 22 0 22 00 0 0 22 0 22 00 0 0 coscos sinsin pr r prr r C pr p prr pr B pr p prr pr A + = + = + + \u2212= + + \u2212= + + = + + = \u03b8\u03b8 \u03b8 \u03d5\u03b8 \u03b8 \u03d5\u03b8\u03b8 \u03d5\u03b8 \u03b8 \u03d5\u03b8\u03b8 (9) Suppose point m on \u03a3 arrives at the contact position after \u03a3 revolves 1\u03b5 in t\u2206 (See Fig.7). At the same time the angle \u03a3\u2032 revolves is 1212 \u03b5\u03b5 i= , and the relative velocity of \u03a3 to \u03a3\u2032 is 12v . The vector 0N reaches the new position 1N : [ ]           \u2212 \u2212 =++= 0 1010 1010 1111 cossin sincos C BA BA kjikCjBiAN \u03b5\u03b5 \u03b5\u03b5 (10) \u03a3 and \u03a3\u2032 satisfy the meshing equation: WVU =\u2212 11 sincos \u03b5\u03b5 (11) Where      \u2212= \u2212= \u2212= xByAW izBiyCV izAixCU 00 210210 210210 (12) Suppose ( )[ ] [ ] ( ) ( )\u03b3\u03d5\u03b8 \u03d5\u03b8\u03d5\u03b8 \u03d5\u03b8\u03d5\u03b8 \u03b3 \u2212+= ++\u2212 +\u2212+ = tg tg tg tg 2 0 22 0 22 0 rpr pr (13) Where 2 0 22 0tg r pr \u03d5\u03b8 \u03b3 \u2212 = (14) Then we have ( ) ( )222 0 4 021 0 221cos \u03d5\u03b8 \u03b4\u03b5 prri pr VU W \u2212+ = + =+ (15) The equation for the contact path is: ( ) ( ) ( ) ( )      = ++\u2212++=+= +++++=\u2212= \u03d5 \u03b5\u03d5\u03b8\u03b8\u03b5\u03d5\u03b8\u03b5\u03b5 \u03b5\u03d5\u03b8\u03b8\u03b5\u03d5\u03b8\u03b5\u03b5 pz rryxy rryxx c c c 101011 101011 cossincossin sincossincos (16) For the spiral involute, we have ck = , \u03b8\u03d5 \u22c5= kc and crp 0= , so 0tg 2 0 22 0 =\u2212= \u2212 = \u03b8\u03b8 \u03d5\u03b8 \u03b3 r pr i"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.37-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.37-1.png",
        "caption": "Figure 2.37. Phasor diagram (a) and transient model (b) of the synchronous generator assuming",
        "texts": [
            "143) T 0 q dE0 d dt \u00fe E0 d \u00bc Xq X0 d Xq Vd (2.15500) M dv dt \u00fe Dv ffi Pm Pe dd dt \u00bc v0v (2.10) or M v0 d2d dt2 \u00fe D v0 dd dt ffi Pm Pe (2.100) After obtaining, for each synchronous generator, the values of E 0 q and E 0 d , at the end of an integration step, the computation is continued by determining a new operating point of the system, followed by a new integration step, and so on. Under these considerations (X0 q \u00bc X0 d), the expression (2.157) of the active power of synchronous generator becomes Pe \u00bc E0 qVd E0 dVq X0 q (2.160) Figure 2.37 illustrates the phasor diagram of voltages and currents assuming X 0 q \u00bc X 0 d and the transient model of the synchronous generator. When the rotor speed changes, the angle d0 of the phasor _E 0 relative to the synchronous reference system (\u00fe1, \u00fej) may be used instead of angle d to measure the change in rotor position. X 0 q \u00bc X 0 d. Therefore, the classical model used for transient stability studies consists of two electromechanical differential equations and one phasor equation as follows: M v0 d2d0 dt2 \u00fe D v0 dd0 dt ffi Pm Pe (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.21-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.21-1.png",
        "caption": "Figure 3.21 The aircraft body axes, (i, j, k), relative to the wind axes, (iv, jv, kv)",
        "texts": [
            " For a rigid aircraft, the aerodynamic and propulsive forces, (X, Y, Z), and moments, (L, M, N), are functions of the relative linear velocity vector, that is, the airspeed, v, as well as the relative flow angles \u2013 the angle of attack, \u03b1, and the sideslip angle, \u03b2, arising out of the orientation of the aircraft\u2019s body axes relative to the velocity vector (or in other words, the linear velocity vector resolved in the body axes).3 We shall employ the relative flow angles, \u03b1, \u03b2, and the geometric bank angle, \u03c3, as Euler angles in order to derive a coordinate transformation between the wind axes, (iv, jv, kv), and the body axes, (i, j, k), as shown in Figure 3.21. Recall from Figure 3.3 that the plane (iv, kv) is the vertical plane, and that (i, k) is the plane of symmetry (the plane containing the lift vector). Therefore, the bank angle, \u03c3, is merely the angle between the two planes (iv, kv) and (i, k) or between the axes jv and j (Figure 3.21). Clearly, the Euler sequence for the necessary coordinate transformation is (\u03b2)3, (\u2212\u03b1)2, (\u03c3)1 (Figure 3.21), resulting in \u23a7\u23a8 \u23a9 iv jv kv \u23ab\u23ac \u23ad = Cv \u23a7\u23a8 \u23a9 i j k \u23ab\u23ac \u23ad , (3.129) where Cv is the following rotation matrix: 3 The vector (X, Y, Z, L, M, N) also depends upon the relative angular velocity vector, (P, Q, R). Cv = \u239b \u239d cos \u03b1 cos \u03b2 cos \u03b1 sin \u03b2 sin \u03b1 \u2212 sin \u03c3 sin \u03b1 cos \u03b2 \u2212 cos \u03c3 sin \u03b2 \u2212 sin \u03c3 sin \u03b1 sin \u03b2 + cos \u03c3 cos \u03b2 sin \u03c3 cos \u03b1 \u2212 cos \u03c3 sin \u03b1 cos \u03b2 + sin \u03c3 sin \u03b2 \u2212 cos \u03c3 sin \u03b1 sin \u03b2 \u2212 sin \u03c3 cos \u03b2 cos \u03c3 cos \u03b1 \u239e \u23a0 . Since the relative velocity vector can be resolved in the body axes as v = v(cos \u03b1 cos \u03b2i + cos \u03b1 sin \u03b2j + sin \u03b1k) (3.130) (Figure 3.21), we have the following relationships among the flow angles, (\u03b1, \u03b2), and the relative velocity components, (U, V, W) (equation (3.119)): \u03b1 = sin\u22121 W\u221a U2 + V 2 + W2 , (3.131) \u03b2 = tan\u22121 V U . (3.132) In a coordinated flight (\u03b2 = 0), with a typically small angle of attack, we have Cv \u239b \u239d 1 0 \u03b1 \u2212\u03b1 sin \u03c3 cos \u03c3 sin \u03c3 \u2212\u03b1 cos \u03c3 \u2212 sin \u03c3 cos \u03c3 \u239e \u23a0 (3.133) or iv i + \u03b1k, jv \u2212\u03b1i sin \u03c3 + j cos \u03c3 + k sin \u03c3, (3.134) kv \u2212\u03b1i cos \u03c3 \u2212 j sin \u03c3 + k cos \u03c3. Flow of atmospheric gases relative to the body\u2019s external surface gives rise to aerodynamic force, fv(t), and torque, \u03c4(t), the fundamental sources of both of which are the surface distributions of static pressure, p(x, y, z, t), acting normal to the surface, and shear stress, f (x, y, z, t), along the relative flow direction:4 fv = \u222b A {\u2212pn + f (n \u00d7 [\u2212iv \u00d7 n])} dA, (3",
            " The elevator can be mounted either forward or behind (oy), usually on the trailing 5 If there is a freestream rotation (swirl), such as that due to a large propeller, or tip-vortices of a large preceding aircraft, it must also be taken into account in equation (3.138). 6 When the vehicle\u2019s structural flexibility is taken into account, it is necessary to include the unsteady aerodynamic effects of control surface deflections, that is, dependence of aerodynamic forces and moments on deflection rates, \u03b4\u0307E, \u03b4\u0307A, \u03b4\u0307R. edge of a larger stabilizing surface in the (xy) plane, as shown in Figure 3.21. Elevator deflection, \u03b4E, creates a forward force, X\u03b4\u03b4E, downforce, Z\u03b4\u03b4E, and a control pitching moment, M\u03b4\u03b4E. A pair of control surfaces in the (xy) plane, located symmetrically about the axis (ox) and deflected in mutually opposite directions by angles \u03b41 and \u03b42 (Figure 3.23) are called ailerons that are used as roll control devices. The aileron deflection is the average of the two separate deflections, \u03b4A = (\u03b41 + \u03b42)/2, and is designed such that a control rolling moment, LA\u03b4A, is produced along with a much smaller, undesirable yawing moment, NA\u03b4A"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002643_amr.308-310.1714-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002643_amr.308-310.1714-Figure9-1.png",
        "caption": "Fig. 9: 1/4 FE model of the tool, collet and clamp ring",
        "texts": [
            " The needed minimum pressure acting on the outer surface of the collet is 194.45MPa and the outer radius of the collet will decrease by 4.87\u00b5m according to Eq. 9. A centrifugal force seriously affects the clamping force when the rotational speed exceeds 1 000 000 rpm [5], so it was ignored here. To clamp the micro-tool, the interference amount between the collet and the clamp ring is at least 53.6\u00b5m according to Eq. 9 and Eq. 10. The effect of the slots was not considered and hence the structure parameters relatively safer. Fig. 9 shows 1/4 FE models of the tool, the collet and the clamp ring with solid185 element type. The interference amount between the collet and the clamp ring is 60\u00b5m and the clearance amount between the tool and the collet 10\u00b5m. The large concentration stress occurs in the tool withdrawal groove (Fig. 10), but which can improve the clamping performance and the stress is within allowable strength. The averaged contact pressure acting on the outside surface of the collet (Fig. 11) is 236.49MPa, which meets the clamping requirement"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003778_978-3-642-17234-2_4-Figure4.18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003778_978-3-642-17234-2_4-Figure4.18-1.png",
        "caption": "Fig. 4.18 The gyrocompass.",
        "texts": [
            "17 Choice of the coordinate systems in order to study the motion of a rigid body relative to a non-inertial frame. lying permanently in the horizontal plane. Without reducing the generality of the problem, we presume that the origins of the frames S; S0; S00 coincide with the fixed point. Next, we take the frame S invariable with respect to the Earth, with its x-axis pointing North and its y-axis pointing West. The x0-axis is taken along the gyroscope axis, while Euler\u2019s angles u; h are chosen as shown in Fig. 4.18. Since the x0-axis remains permanently in the horizontal plane, we have w = 0. It is more convenient to express all quantities relative to the frame S0 (e.g. the ship, or the airplane where the gyrocompass is installed). In this frame, the vector x \u00bc _uk\u00fe _hi0 has the components: x01 \u00bc _h; x02 \u00bc 0; x03 \u00bc _u: \u00f04:6:42\u00de Since Ox0 is a principal axis of inertia, we have I02 \u00bc I03 6\u00bc I01: The angular momentum and the kinetic energy of the gyroscope are given by (4.5.28) and (4.5.29), respectively: L01 \u00bc I01 _h; L02 \u00bc 0; L03 \u00bc I03 _u; \u00f04:6:43\u00de T \u00bc 1 2 I01 _h2 \u00fe 1 2 I02 _u2: \u00f04:6:44\u00de Let c be the angle between the z-axis and the instantaneous vector of rotation X of the Earth (Fig. 4.18). Then, X01 \u00bc X0 sin c cos u; X02 \u00bc X0 sin c sin u; X03 \u00bc X0 cos c; such that I0ab X0a X0b \u00bc I02 X02 cos2 c\u00fe X02 sin2 c\u00f0I01 cos2 u\u00fe I02 sin2 u\u00de; X Lrot \u00bc I01 _h X0 sin c cos u\u00fe I02 _u X0 cos c: Therefore, the corresponding Lagrangian is L \u00bc 1 2 I01 _h2 \u00fe 1 2 I02 _u2 \u00fe 1 2 X02 sin2 c\u00f0I01 cos2 u\u00fe I02 sin2 u\u00de \u00fe I02 X02 cos2 c\u00fe I01 _hX0 sin c cos u\u00fe I02 _uX0 cos c; \u00f04:6:45\u00de where the constant potential energy has been omitted. We observe that the generalized coordinate h is cyclic, so that we have the first integral ph \u00bc oL o _h \u00bc I01\u00f0 _h\u00fe X0 sin c cos u\u00de \u00bc C1: \u00f04:6:46\u00de Another first integral follows from the fact that the energy is conserved, since the Lagrangian L does not depend explicitly on time"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002643_amr.308-310.1714-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002643_amr.308-310.1714-Figure5-1.png",
        "caption": "Fig. 5: Stress caused by axial force Fig. 6: Stress caused by deviation",
        "texts": [
            " 2 shows the maximum positive stress on the elastic shaft coupling as a function of the height and length of the thin wall according to Eq. 6. All kinds of height with corresponding length can meet strength condition. The height of 0.4mm and length of 2mm were primarily chosen. A 1/4 finite element (FE) model (Fig. 3) of the elastic shaft coupling with solid45 element type was founded in a finite element package ANSYS and the FE analysis was conducted. The maximum Von Mises stress caused by deviation of 0.1mm is about 361.226MPa (Fig. 6), which is far more than that by torque of 5N\u00b7mm (Fig. 4) and that by axial force of 5N (Fig. 5). The maximum stress caused by the combined load (Fig. 7) is about 366.132MPa, which is less than allowable stress and close to that by deviation and so the strength of the elastic shaft coupling is mainly determined by deviation amount. Fig. 8 shows changes of stresses caused by different loads in the course of rotation. The stress caused by torque and that by axial force remain unchanged, but the maximum stress caused by deviation and that by combined load vary cyclically with the rotating of the shaft"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-FigureB-1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-FigureB-1-1.png",
        "caption": "Figure B-1. Output Angle per Tooth Hit on the Planet Gear",
        "texts": [
            " Similarly, the rotational angle \u03b8BP on planet gear \u201cB\u201d of the compound planet gear can be derived as: ])([360 N Ni N Niround N BRBR BP BP \u22c5 \u2212 \u22c5 \u22c5= o \u03b8 i =1 to N (A7) The clocking angle of the ith planet gear \u201cB\u201d relative to planet gear \u201cA\u201d is: APBPiclocking \u03b8\u03b8 \u2212=_ i =1 to N (A8) ])([360])([360 N Ni N Niround NN Ni N Niround N ARAR AP BRBR BP \u22c5 \u2212 \u22c5 \u22c5\u2212 \u22c5 \u2212 \u22c5 \u22c5= oo where a positive result means it is in the CCW direction, and a negative result means it is in the CW direction for equally spaced planet gears. Derivation of Tooth Hits on Simple Planetary Gears A simple planetary gear set is shown in Figure 22. The sun gear is the input, and meshes with planet gears. There are a total of N planet gears. The ring gear which is fixed to the ground meshes with planet gears, too. The carrier is the output. For planet gears (Figure B-1), they will get hit once when \u03b8C + (-\u03b8P) = 360\u25cb (B1) From Table 1, the following rotational relationship can be derived as: R S P S R S C P N N N N N N \u2212 = \u03b8 \u03b8 (B2) After substituting Equation (B2) into (B1), the output angle per tooth hit on the planet gear can be derived as: \u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b = R P C N N 1 360o\u03b8 (B3) For the sun gear (Figure B-2), it will get hit once when \u03b8S - \u03b8C = 360\u25cb/N (B4) From Table 1, the following rotational relationship can be derived as: R S R S C S N N N N 1+ = \u03b8 \u03b8 (B5) After substituting Equation (B5) into (B4), the output angle per tooth hit on the sun gear can be derived as: Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003431_2013-36-0272-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003431_2013-36-0272-Figure4-1.png",
        "caption": "Figure 4 \u2013 Tire belt segment",
        "texts": [
            " We will also suppose that the torque applied on the wheel rim will be entirely transmitted to the tire belt through the sidewalls, and then from the tire belt to the contact region, neglecting the small portion transmitted directly to the ground by the sidewalls. This is an approximated model for current tires, but it is an exact representation of the behavior of \u201ctweels\u201d (tire + wheels) proposed by Michelin [1], [3], illustrated in Figure 3. Next, we will consider the effect of the tire belt\u2019s deformation. Let us take a segment AB of the tire belt, as shown in Figure 4, with length l0 and mass m, resulting in a linear mass density given by: (2) If this segment is deformed (stretched) by some force F, its new length will be l and the deformation is given by: from which we may write: (3) The total mass m of this segment does not change, but the linear mass density changes to , given by: Page 3 of 7 Using equations (2) and (3), we may write: (4) Let us suppose now that we have a moving wheel, and we define a translational relative referential (O, x*, y*) fixed to the center O, moving to the right with velocity VO, as shown in Figure 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001687_978-3-7091-1187-1_5-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001687_978-3-7091-1187-1_5-Figure2-1.png",
        "caption": "Figure 2. Bilinear uniaxial stress-strain curve. In mixed hardening the yield stress in compression, \u2212\u03c3C , is smaller than \u03c3B \u2212 2\u03c3A and larger than \u2212\u03c3B .",
        "texts": [
            " It is as well necessary to re-calculate the generalized displacements. These are the tasks of a second cycle. The procedure here suggested to compute the plastic strains is the governing parameter method with the increment of effective plastic strain, \u0394eP , as governing parameter (Kojic\u0301 and Bathe (2005)). This procedure is described in the following paragraphs. The von Mises yield function, our governing function, cannot be greater than zero. A bilinear stress-strain relation with mixed hardening is here assumed, Figure 2, and, therefore, the yield function can be written as fy ( \u0394eP ) = t+\u0394t\u03c3E t+\u0394t\u03c3y + [3G+ (1\u2212M)EP ] \u0394eP \u2212 1 (31) When yielding takes place, \u0394eP is calculated such that the governing function fy ( \u0394eP ) is zero. M in equation (31) is a mixed hardening parameter, which is a characteristic of the material and quantifies Bauschinger effect (Dieter (1986)). A mixed hardening material model (0< M <1) is placed between isotropic (M=0) and kinematic hardening material models (M=1). Only the isotropic part of the effective plastic stress affects the size of the yield surface (Kojic\u0301 and Bathe (2005))"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002238_amr.476-478.579-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002238_amr.476-478.579-Figure3-1.png",
        "caption": "Fig. 3 Stress cloud charts of short flax reinforced polypropylene composites at different time",
        "texts": [
            " Stress-strain curve has been achieved and analyzed. Between the MSC.Patran/Nastran and numerical simulation of polypropylene materials, setting of geometry model, grid partition, working and boundary condition, unit attributes and analysis were the same. It is observed from Fig. 2 that stress-strain curves of thermoplastic composites obtained by changing fiber volume fraction. It can be seen that modulus of the composites increases with the rise in fiber volume fraction and at initial stage, approximately linearly. Fig. 3 shows stress cloud charts of short flax fiber-reinforced polypropylene composites at different time. The fiber volume content was 30%.From Fig. 3, it is clear that loading area changes constantly when loading time increasing and the deformation (compression) has been transformed into tension. Effect of Fiber orientation distribution Due to different fabricate methods of composite webs and felts, short fiber orientation distribution in composites were not the same. In this paper, flexural property of carding thermo-Bonded, carding needle-punching samples, gilling thermo-Bonded, gilling needle-punching samples, laminated thermo-Bonded and laminated needle-punching samples was numerical evaluated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.9-1.png",
        "caption": "Fig. 3.9 Dual string pendulum. (a) In rest position, view from front. (b) Rigid massless rod rotated through angle , view from top",
        "texts": [
            "10) where is the radius of gyration of the square, and t is time. The approximate time of swing, T , is T D 2 s 2l gc2 (3.11) which differs from Eq. 2.13 by the factor =c. A dual string pendulum is a simplification of the quadrifilar pendulum (previous section), and its motions are analogous. It consists of two identical simple string pendulums (Sect. 2.4) with the point masses connected by a rigid massless rod, length 2c. In the rest position the strings are vertical and the rigid massless rod is horizontal, as shown in Fig. 3.9a. Provided that the strings remain taut a dual string pendulum has three degrees of freedom. If the rigid massless rod remains parallel to its rest position then there are two degrees of freedom, and the motion of each point mass, m, is in phase and identical and also, in general, identical to the motion of the point mass, m, of a simple string pendulum (Sect. 2.4). Exceptions are due to inference between components of the dual string pendulum. The third degree of freedom means that a torsional mode of oscillation is possible, in which the rigid massless rod rotates about a vertical axis (Fig.3.9b). In this torsional mode of oscillation the simple string pendulums are moving out of phase, that is the phase is 180\u0131. For the dual string pendulum the radius of gyration, , of the rod and point masses about the vertical axis through its centre is c, so Eq. 3.11 reduces to Eq. 2.13. Hence, for small amplitudes the time of swing is independent of the length of the rigid massless rod. In mechanical engineering terms the trapezium pendulum is a Watt\u2019s linkage, developed in 1784 (Tre\u0301baol 2008), and sometimes called Watt\u2019s parallel motion (Dunkerley 1910)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003350_icma.2012.6284390-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003350_icma.2012.6284390-Figure1-1.png",
        "caption": "Fig. 1 move frame of vessel",
        "texts": [
            " In the line motion, the heave motion is the greatest impact factor of vessel performance. Body frame(b) and geographic frame(t) are usually used in strap-down inertial navigation system. In this paper, move frame(d) is introduction to describe the heave movement. Move frame is moved with vessel speed but not rotated with vessel [10,11]. Its position is coincided with body frame when the vessel movement is not six degree of freedom. The relationship between the geographic frame and move frame is shown in figure 1. Where G\u03c8 is main heading angel, \u03b7 is yawing angel, move frame has relationship with the main heading angel, and not change with six degree of freedom of movement. The move frame is given by geographic frame rotate main heading angel around z-axis. The vessel reciprocating motion in a certain periodic along with z-axis of move frame is occurred with the effect of wave. This motion is called as heave motion. III. THE MEASUREMENT SCHEME OF HEAVE MOTION Heave movement is an information of displacement, and its measurement must take advantage of the accelerometer of strap-down inertial navigation system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003528_amr.706-708.1209-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003528_amr.706-708.1209-Figure2-1.png",
        "caption": "Fig. 2 Tangential error characteristics",
        "texts": [
            " Features of grinding involute gears on a worm wheel gear grinding machine Usually, there are serious tooth-period errors in the process of grinding Involute gears on a worm wheel gear grinding machine. And tooth-period errors consist of radial errors caused by centre\u2013to\u2013centre spacing of the grindstone and the machined gear( such as a radial runout of grindstone, and lack of dynamic balance), tangential errors caused by circumferential position errors between the grindstone and the work-piece( such as transmission errors, a axial inching, helix errors of grindstone, etc). Referencing Fig. 1(a) and Fig. 2(a), the axial section can be treated as a rack. The process of grinding a gear with grindstone is regarded as the engaging process of the machined gear and the rack[3]. (1) In the initial position, the rack (grindstone) is tangent to machined gear tooth at point Q and point 'Q . The two points are just right at constant chord, which are considered as the start points of tooth profile error computation. (2) When the machined gear turns, the contact points on left and right tooth surface take the rightabout movement from point Q and 'Q ",
            " Characteristics of radial error: (a) Profile errors of corresponding point of left and right tooth surface are equal. ' rr tt \u2206=\u2206 . (b) The wavelength of profile error curve is the base pitch. (c) With different tool setting position, the phase position of profile error curve shows signs of change. Characteristics of tangential errors of a worm wheel gear grinding machine When the grindstone has tangential errors, left and right tooth profile errors with equal value and opposite direction are caused (referencing Fig. 2(a). ' tt tt \u2206\u2212=\u2206 . Referencing Fig. 2(b), the error of point d\u2019 is positive, the error of point d is negative, the error of point e\u2019 is positive, the error of point e is negative, left and right tooth profile error curves show accordant signs. If the frequency of tangential error is f, then profile errors show as the following: )sin( ttt fet \u03d5\u03d5 +=\u2206 , )sin(' ttt fet \u03d5\u03d5 +\u2212=\u2206 (2) If tangential errors are caused by a axial inching and helix errors of grindstone, then f = z. )sin( ttt zet \u03d5\u03d5 +=\u2206 , )sin(' ttt zet \u03d5\u03d5 +\u2212=\u2206 (3) Thereinto, te is a tangential eccentric, z is gear tooth number, \u03d5 is the gear turning angle, t\u03d5 is the initial angle. If the spread length is thought as the longitudinal coordinate of the gear profile error curve, then it is easy to see that the gear profile error caused by a tangential eccentric has a period of each grindstone turn and the wavelength is the base pitch. And with different tool setting position, the phase position of profile error curve shows signs of change. For example, referencing Fig. 2(b), there is a axial inching to the left bone at \u00b00 , and there is a axial inching to the right bone at \u00b0180 , then ed tt \u2206\u2212=\u2206 ' , 'ed tt \u2206\u2212=\u2206 . And so do in other tool setting spots, referencing Fig. 2(c), (d). (e). Characteristics of tangential error: (a) Profile errors of corresponding point of left and right tooth surface show a sign of equal value and opposite direction. ' tt tt \u2206\u2212=\u2206 . (b) In majority situation, the wavelength of profile error curve is the base pitch. (c) With different tool setting position, the phase position of profile error curve shows signs of change. Combination of a radial error and a tangential error As the matter of fact, radial errors and tangential errors coexist in a worm wheel gear grinding machine"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.13-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.13-1.png",
        "caption": "Figure 12.13 A vehicle with roll and yaw rotations.",
        "texts": [
            "254) Substituting the associated vectors generates the Newton equations of motion in the body coordinate frame: \u23a1 \u23a3 Fx Fy 0 \u23a4 \u23a6 = m GRT B \u23a1 \u23a2 \u23a3 ( v\u0307x \u2212 \u03c8\u0307 vy ) cos \u03c8 \u2212 (v\u0307y + \u03c8\u0307 vx ) sin \u03c8( v\u0307y + \u03c8\u0307 vx ) cos \u03c8 + (v\u0307x \u2212 \u03c8\u0307 vy ) sin \u03c8 0 \u23a4 \u23a5 \u23a6 = m \u23a1 \u23a3 v\u0307x \u2212 \u03c8\u0307 vy v\u0307y + \u03c8\u0307 vx 0 \u23a4 \u23a6 (12.255) Applying the same procedure for moment transformation, GMC = GRB BMC \u23a1 \u23a3 0 0 MZ \u23a4 \u23a6 = \u23a1 \u23a3 cos \u03c8 \u2212 sin \u03c8 0 sin \u03c8 cos \u03c8 0 0 0 1 \u23a4 \u23a6 \u23a1 \u23a3 0 0 Mz \u23a4 \u23a6 = \u23a1 \u23a3 0 0 Mz \u23a4 \u23a6 (12.256) we find the Euler equation in the body coordinate frame: Mz = \u03c9\u0307z Iz (12.257) Example 736 The Roll Model Vehicle Dynamics Figure 12.13 illustrates a vehicle with a body coordinate frame B(Cxyz) at the mass center C. The x-axis is a longitudinal axis passing through C and directed forward. The y-axis goes laterally to the left from the driver\u2019s viewpoint. The z-axis makes the coordinate system a right-hand triad. When the car is parked on a flat horizontal road, the z-axis is perpendicular to the ground, opposite to the gravitational acceleration g. The equations of motion of the vehicle are usually expressed in B(Cxyz). The angular orientation and angular velocity of the vehicle are expressed by three angles\u2014roll \u03d5, pitch \u03b8 , and yaw \u03c8 \u2014and their rates\u2014roll rate p, pitch rate q, and yaw rate r: p = \u03d5\u0307 q = \u03b8\u0307 r = \u03c8\u0307 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003534_amr.308-310.2220-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003534_amr.308-310.2220-Figure2-1.png",
        "caption": "Fig. 2 Cycloid gear and output mechanism finite element model",
        "texts": [
            " First using three-dimensional modeling software POR/E establish FA cycloid drive transmission whole assembly model ,and then using the seamless connection interfaces between POR/E and ANSYS, put the built assembly model to the ANSYS, then obtained the geometrical model shown as in the Fig.1. Finite element grid partition Considering various factors may influence on the quality of the grid partition and the geometrical characteristic of cycloid gear and output dowel pin, using the body sweeping grid partition method, and manual control of the grid partition, the finite element mesh model is finished shown as in the Fig. 2. Among them, part a) is the whole finite element model, part b) is the local finite element model, node number is 455756, unit number is 437612. 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: 152.14.136.77, NCSU North Carolina State University, Raleigh, USA-09/05/15,02:44:44) Because the simplified model had set dowel pin bush and output dowel pin as a whole model, take materials of cycloid gear and output dowel pin were all GCr15SiMn"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003294_j.engfailanal.2010.12.022-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003294_j.engfailanal.2010.12.022-Figure1-1.png",
        "caption": "Fig. 1. Failed tappets: (a) overlook view; (b) front view.",
        "texts": [
            " The amount of carbide (M3C) is demanded more than 40% within the chilled depth. Tempered martensite of matrix microstructure should be obtained after quenching and tempering process. The chemical composition of the failed tappets was determined by spectrographic chemical analysis method. The axialsection microstructure was observed by OPM and SEM. Wear morphology on the end of tappets was observed by visual and SEM. The hardness profiles on the chilled end were made by Rockwell meter. Three serviced tappets and a new tappet are shown in Fig. 1 (labeled as Nos. 1, 2, 3 and 4). Very severe wear occurred at the end of tappet 1, whose end had been worn out to form a hole of 10 mm 8 mm in the centre and oxidation color presents on the end (Fig. 1a). The end of tappet 2 became a concave by intensely wear (Fig. 1a). The axial wear thickness at the external edge on the end for tappets 1 and 2 is respectively about 5.0 mm and 2.5 mm, while, only 0.2 mm for tappet 3 . All rights reserved. ; fax: +86 0411 84728670. (Fig. 1b). The longitudinal-sections of tappets (Fig. 2) show that the height difference between the concave point at the centre and the external edge of the end for tappets 1 and 2 is respectively about 1.5 mm and 1.0 mm, but the end of tappet 3 is relatively even. SEM observation shows that rough wear tracks appeared on the end of tappet 1, on which extensive plough and plastic flow [1,2] were exhibited (Fig. 3a). It may be deduced that the end hardness of tappet 1 is lower than that of its friction couple \u2013 the camshaft (PHRC48 as the specified)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.5-1.png",
        "caption": "Fig. 7.5 Getting-busier directions for kinematics and dynamics",
        "texts": [
            " However, it is just a computer recursive numerical algorithm, and though fast, cannot be used for globally modeling a robotic dynamics. With the exponential growth of computer power in the recent years, the Lagrangian formulation has been regaining its popularity. The central idea of the Newton-Euler algorithm is to update each force or torque that acts on each link based on the Newton\u2019s Second and Third Laws plus the D\u2019Alembert\u2019s Principle in classical mechanics. It is observable that by directly inspecting a robot arm, as shown in Figure 7.5, the direction of getting kinematically \u201cbusier\u201d is from frame 0 (the base) forward to frame n (the hand), while the direction of getting dynamically \u201cbusier\u201d is just opposite. Therefore, we should first calculate both the linear and angular velocities as well as the accelerations for each link from link 1 recursively to link n as an outward procedure, and then compute each joint torque or force from joint n recursively back to joint 1 as an inward procedure. Figure 7.6 illustrates a schematic force and torque analysis of link i that is interacted by both two neighboring links i \u2212 1 and i + 1"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001725_9781119971191.ch3-Figure3.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001725_9781119971191.ch3-Figure3.5-1.png",
        "caption": "Figure 3.5 External forces in the vertical plane containing the relative velocity vector",
        "texts": [
            " Thus, the lift and drag are modulated by changing the angle of attack with nearly constant airspeed and constant atmospheric density of a cruising flight. For a conventional aircraft without the capability of thrust vectoring, one can readily approximate \u03b5 \u03b1, which implies that the thrust nearly acts along the longitudinal reference line in the plane of symmetry used for measuring the angle of attack. Thus, for the aircraft navigational plant we shall adopt this convention, which results in the forces in the vertical plane containing the velocity vector, (iv, kv), as shown in Figure 3.5. The atmosphere is always in motion due to combined effects of planetary rotation and thermodynamic variations in local atmospheric temperature and pressure, leading to strong horizontal air currents, called winds. If unaccounted for, winds can cause large errors in the position of airplanes, which can lead to a shortfall in range and fuel starvation. Therefore, careful periodic observations of the wind velocities above selected weather stations along the route, yielding hourly winds-aloft data at a series of altitudes, are essential for accurate navigation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure6-1.png",
        "caption": "Fig. 6 A push-push type connection",
        "texts": [
            " The structure represented above can make a turning motion to connect or disconnect the male and female sides. While designing this kind of mechanical quick connection for medical devices, the angle C (in Fig.5) is the key point to make sure the structure works. A. Stricture Scheme A new kind of quick connector was designed based on the structure of a push-push type latch mechanism and electronic device thereof [9]. This machine was designed for a connection of a function bed and a devise which hold a tubular shape medical device, a catheter [10] for example. Fig.6 shows a full view of the connector. It includes a holding device which is the male part of the connection, and the locking part links with the function bed, which is the female part of the connection. Details of the structure are shown in Fig.7. The process of this connection will be descript as following: firstly, push the holding device down to the base plate, then the side wall of the holding device will press the locking pin, top of the locking pin moves on the surface of the holding device"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.6-1.png",
        "caption": "Fig. 2.6 4PRRRR-type fully-parallel PMs with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology P\\R\\R||R\\R (a) and P||R\\R||R\\R (b)",
        "texts": [
            "1c) The two last revolute joints of the four limbs have parallel axes 4. 4RRRRR (Fig. 2.3b) R||R||R\\R||R (Fig. 2.1c) The three first revolute joints of the four limbs have parallel axes 5. 4RRRRR (Fig. 2.4a) R||R\\R||R||R (Fig. 2.1d) The two first revolute joints of the four limbs have parallel axes. 6. 4PRRRR (Fig. 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7. 4RRRPR (Fig. 2.5a) R||R\\R\\P\\kR (Fig. 2.1f) Idem No. 5 8. 4RRPRR (Fig. 2.5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10. 4PRRRR (Fig. 2.6b) P||R\\R||R\\R (Fig. 2.1i) Idem No. 9 11. 4RRPRR (Fig. 2.7a) R\\R\\P\\kR\\R (Fig. 2.1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2.8a) P||R||R||R||R\\R (Fig. 2.1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig. 2.9a) R||R||P||R\\R (Fig. 2.1n) Idem No. 13 16. 4RRRPR (Fig. 2.9b) R||R||R||P\\R (Fig. 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.11-1.png",
        "caption": "Fig. 3.11 Trapezium pendulum. (a) In rest position. (b) Definition of pendulum angle",
        "texts": [
            " In a four bar linkage the four links are pivoted together so that they move within a plane, as shown schematically in Fig. 3.10, These links are: a static bar, regarded as being fixed in space, a pilot crank, a connecting rod, and a rocker. A coupler may be attached to one of the moving links. A four bar linkage has one degree of freedom since one parameter, for example the angle between the static bar and the pilot crank, is needed to specify its configuration. The trapezium pendulum as a form of Watt\u2019s linkage is shown in Fig. 3.11. The pilot crank and rocker of a four bar linkage are the two rigid massless rods, length l1, shown in the figure. At their lower ends they are connected by horizontal frictionless pivots to a rigid massless rod, length l2, with a point mass, m, at its centre. This is horizontal in the rest position (Fig. 3.11a), and is the connecting rod in a four bar linkage. At their upper ends the two rigid massless rods, length l1, are fixed in space by two fixed horizontal frictionless pivots, at the same level, and a distance l3 apart. The two fixed frictionless pivots correspond to the static bar in a four bar linkage. The complete path of the point mass, m, is a figure of eight with two nearly straight portions, which are exploited in Watt\u2019s parallel motion. Part of the path, shown by a solid line in Fig. 3.11a, is a non circular arc. Design methods for four bar linkages in general (Hrones and Nelson 1951), and Watt\u2019s linkage in particular, are well established (Dunkerley 1910; Tre\u0301baol 2008). However, equations for the motions of four bar linkages are cumbersome, and there is no simple expression for the motion of the point mass, m, in the trapezium pendulum, Hence there is, in general, no simple expression for its time of swing. At the rest position the path of the point mass, m, has radius l4, where l4 is the vertical distance from the fixed frictionless pivots to the point mass (Fig. 3.11a). Hence, from Eq. 2.13 for a simple rod pendulum, the time of swing, T , for small amplitudes is given approximately by T D 2 s l4 g (3.12) where g is the acceleration due to gravity. The path that would be followed by the point mass, m, of a simple rod pendulum, length l4. is shown by the circular arc in Fig. 3.11a. For the same pendulum angle, , (Figs. 2.2 and 3.11b) the path of the point mass, m, of the trapezium pendulum deviates increasingly from the circular arc as the pendulum angle, , increases. In consequence, the time of swing, T , of the trapezium pendulum decreases relative to that of a simple rod pendulum. In the special case where l2 D l3 the pilot crank and rocker of a four bar linkage are parallel, and a trapezium pendulum becomes a dual rod pendulum. A dual rod pendulum consists of two rigid massless rods, length l , suspended from frictionless pivots connected to a horizontal rigid massless rod with a point mass, m, at its centre"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.7-1.png",
        "caption": "Fig. 7.7 Velocity analysis of a three-joint planar robot arm",
        "texts": [
            " The first approach, though straightforward, works only in relatively simple cases, such as a planar robot or a 3D robot with up to 3 joints, because a robot with more links will exponentially increase the complexity of its velocity analysis and difficulty of geometric inspection. In contrast, the second approach will offer a global solution that allows us to determine a larger inertial matrixW both symbolically and numerically for most types of robotic systems and even for a large-scale digital human model [4, 12, 13]. Let us now look at a three-joint planar robot arm, as shown in Figure 7.7, to illustrate the first approach to finding its 3 by 3 inertial matrix W via a total kinetic energy derivation. This planar robot consists of two revolute joints and one prismatic joint as an RPR type robot arm. Let q = (\u03b81 d2 \u03b83) T be the joint position vector. Each of the three links has its mass mi and the moment of inertia Ii about the axis passing through the mass center of each link and perpendicular to the plane for i = 1, 2, 3. In Figure 7.7, each lci is the length between the mass center of link i and the origin of frame i. Based on equation (7.11), the kinetic energy K1 of link 1 can be written as K1 = 1 2 m1l 2 c1\u03b8\u0307 2 1 + 1 2 I1\u03b8\u0307 2 1, because the mass center velocity is given by v = \u03c9r = lc1\u03b8\u03071. Since the motion of link 2 is imposed by the rotation of link 1 in addition to its own sliding with a linear velocity d\u03072, the kinetic energy K2 of link 2 becomes K2 = 1 2 m2[(d2 \u2212 lc2)2\u03b8\u030721 + d\u030722] + 1 2 I2\u03b8\u0307 2 1 . Note that the two velocities (d2 \u2212 lc2)\u03b8\u03071 and d\u03072 for link 2 happen to be perpendicular to each other so that the resultant velocity square is equal to the sum of the two velocity squares. If they are not perpendicular, we have to project each of them onto a local orthogonal frame originated at the common acting point and then sum up their squares. For link 3, there are three linear velocities: the self-rotation linear velocity lc3(\u03b8\u03071 + \u03b8\u03073) because the total angular velocity of link 3 is now \u03b8\u03071 + \u03b8\u03073, plus two imposed velocities d2\u03b8\u03071 and d\u03072 by the motion of link 2, which are all delivered to the mass center of link 3, see Figure 7.7. Since the last two imposed velocity vectors are perpendicular to each other, we may utilize them as the local orthogonal frame just for the purpose of projection convenience. Thus, we have K3 = 1 2 m3[(d2\u03b8\u03071 + lc3(\u03b8\u03071 + \u03b8\u03073)c3) 2 + (d\u03072 \u2212 lc3(\u03b8\u03071 + \u03b8\u03073)s3) 2] + 1 2 I3(\u03b8\u03071 + \u03b8\u03073) 2 = 1 2 m3[d 2 2\u03b8\u0307 2 1 + d\u030722 + 2d2lc3\u03b8\u03071(\u03b8\u03071 + \u03b8\u03073)c3 \u2212 2lc3d\u03072(\u03b8\u03071 + \u03b8\u03073)s3 + +l2c3(\u03b8\u03071 + \u03b8\u03073) 2] + 1 2 I3(\u03b8\u0307 2 1 + \u03b8\u030723 + 2\u03b8\u03071\u03b8\u03073), where c3 = cos \u03b83 and s3 = sin \u03b83. Once a symbolical form of the total kinetic energy K = K1 +K2 +K3 is achieved, the 3 by 3 inertial matrix W can directly be found by extracting and augmenting all the coefficients from K",
            " The Riemannian metricW32 on the minimum embeddable C-manifold can be found by W32 = JTJ = \u239b \u239d w11 w12 w13 w12 b2 + 2bdc3 + d2 + h2 bdc3 + d2 + h2 w13 bdc3 + d2 + h2 d2 + h2 \u239e \u23a0 , where w11 = a2 + b2 + 2abc2 + 2adc23 + 2bdc3 + d2 + h2 + h21, and w12 = b2 + abc2 + adc23 + 2bdc3 + d2 + h2, w13 = adc23 + bdc3 + d2 + h2. One can also verify without difficulty that by adjusting all the parameters a, b, d, h and h1, the metric W32 = JTJ can be equal to the inertial matrix W of this RRR-type planar robot. Such a parameter adjustment just plays the role of deformer D(\u03a8) in the C-manifold isometrization. In contrast to the above RRR-type planar robot arm, let us revisit the RPR-type planar robot, as given in Figure 7.7 or 7.9. At first glance, it seems sufficient enough to pick up only the last link to represent the minimum embeddable C-manifold for this particular robot, because the configuration becomes unique whenever the last link is motionless. If so in this case, n = 3 and k = 1. Based on the previous kinematics analysis of this planar robot, the tip position vector and the position vector of frame 2 can be found as pt0 = \u239b \u239d d2s1 + d4s13 \u2212d2c1 \u2212 d4c13 0 \u239e \u23a0 , and p20 = \u239b \u239d d2s1 \u2212d2c1 0 \u239e \u23a0 where d4 is the length of the last link"
        ],
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    },
    {
        "image_filename": "designv11_100_0003610_imece2013-63166-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003610_imece2013-63166-Figure3-1.png",
        "caption": "FIGURE 3: VELOCITY VECTORS FOR HORIZONTAL AND VERTICAL CONFIGURATION",
        "texts": [
            " Case 1: \u03c6 = atan ( (c\u03c8 \u2212 c\u03b8)A\u2212 s\u03b8s\u03c8B c\u03b8C\u2212 s\u03c8s\u03b8A\u2212 c\u03c8B ) (2) where A = c\u03b1 \u2212 c\u03b2 , B = s\u03b1 \u2212 s\u03b2 , C = (c\u03b2 \u2212 1)/tan\u03b2 \u2212 (c\u03b1 \u22121)/tan\u03b1 y =\u2212rs\u03c6c\u03b8 (3) x =\u2212c\u03c6c\u03b8c\u03b1r\u2212 (s\u03c6c\u03c8 + c\u03c6s\u03b8s\u03c8)s\u03b1r+ (4) + r (s\u03c6c\u03b8 (c\u03b1 \u22121)+(c\u03c6c\u03c8 + s\u03c6s\u03b8s\u03c8)s\u03b1) tan\u03b1 Case 2: \u03c6 = atan ( A(c\u03c8 \u2212 c\u03b8)\u2212 s\u03c8s\u03b8C c\u03b8C\u2212 s\u03c8s\u03b8A\u2212 c\u03c8B ) (5) y =\u2212r (s\u03c6c\u03b8 + c\u03c6s\u03c8s\u03b8) (6) x =\u2212(c\u03c6c\u03b8 \u2212 s\u03c6s\u03c8s\u03b8)rc\u03b1 + s\u03c6c\u03c8rs\u03b1+ (7) + r ((s\u03c6c\u03b8 + c\u03c6s\u03c8s\u03b8)(c\u03b1 \u22121)+ c\u03c6c\u03c8s\u03b1) tan\u03b1 Notice, there is no dependency on z in any of the three equations that define the parasitic motions. However, all the parasitic motions depend on the rotations \u03c8 and \u03b8 , and the constant values \u03b1 and \u03b2 . If the inverse position problem is solved, the expressions of the actuated joint variables are obtained, as shown in Eqn. (8): bi = rixc\u03c4 + riys\u03c4 \u00b1 \u221a l2 i \u2212 (rixs\u03c4 \u2212 riyc\u03c4)2 \u2212 r2 iz (8) where \u03c41 = 0, \u03c42 = \u03b1 and \u03c43 = \u03b2 . Differentiating Eqn. (1), the vectorial expression of the velocity of the mobile platform is obtained, Eqn. (9) (see Fig. 3). vp = b\u0307iki +\u03c9i \u00d7 li \u2212\u03c9p \u00d7 rAi (9) This equation can be expressed using matrix notation, Eqn. (10): Jx [ vp \u03c9p ] = Jq b\u03071 b\u03072 b\u03073  (10) where Jx is the inverse Jacobian matrix and Jq is the direct Jacobian matrix. 3 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/02/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Again, if Eqn. (9) is differentiated, the vectorial expression of the acceleration is obtained, Eq. (11): ap = b\u0308iki +\u03b1i \u00d7 li +\u03c9i \u00d7 (li \u00d7\u03c9i)\u2212\u03b1p \u00d7 rAi +\u03c9p \u00d7 (rAi \u00d7\u03c9p) (11) In matrix notation, Eqn"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.38-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.38-1.png",
        "caption": "Figure 2.38. General dynamic model of the electromagnetic part of a synchronous generator for",
        "texts": [
            " Under these assumptions, the operational equations with external parameters that give the synchronous generator performances in transient operation, expressed in the coordinates system (d, q) solidary with the rotor, are components of the terminal voltage: Vd ffi vcq; Vq ffi vcd components of stator flux linkages: cd \u00bc A s\u00f0 \u00deVf Ld s\u00f0 \u00deId (2.87) cq \u00bc Lq s\u00f0 \u00deIq (2.79) where s denotes the operator d/dt. We observe from (2.87) that the d-axis component of the stator flux is dynamically related to the field voltage and to the d-axis component of the armature current. Also, from (2.79) we see that the q-axis component of the stator flux is related only to the q-axis component of the armature current. Generic dynamic models (Figure 2.38) can be drawn based on equations (2.87) and (2.79) [12]. In the sixth-order dynamic model there are a field circuit and an additional rotor circuit that acts along the d-axis and two additional rotor circuits that act along the q-axis. The transfer functions Ld(s), Lq(s), and A(s) are expressed under the form [1,2,12,29] Ld s\u00f0 \u00de \u00bc Ld 1\u00fe sT 0 d 1\u00fe sT 00 d 1\u00fe sT 0 d0 1\u00fe sT 00 d0 (2.880) Lq s\u00f0 \u00de \u00bc Lq 1\u00fe sT 0 q 1\u00fe sT 00 q 1\u00fe sT 0 q0 1\u00fe sT 00 q0 (2.92) A s\u00f0 \u00de \u00bc Lmd Rf 1\u00fe sTkd\u00f0 \u00de 1\u00fe sT 0 d0 1\u00fe sT 00 d0 (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.18-1.png",
        "caption": "Fig. 2.18 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRCR (a) and 4RCRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R||R||C\\R (a) and R||C||R\\R (b)",
        "texts": [
            " 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig. 2.1g0) Idem No. 13 34. 4RRRC (Fig. 2.19a) R\\R||R||C (Fig. 2.1h0) Idem No. 21 35. 4RCRR (Fig. 2.19b) R\\C||R||R (Fig. 2.1l0) Idem No. 21 36. 4RRCR (Fig. 2.20a) R\\R||C||R (Fig. 2.1b0) Idem No. 21 37. 4RCRR (Fig. 2.20b) R||C\\R||R (Fig. 2.1c0) Idem No. 21 j\u00bc1 fj 5 5 23. Pp2 j\u00bc1 fj 5 5 (continued) In the fully-parallel topologies of PMs with coupled Sch\u00f6nflies motions F / G1G2-G3-G4 presented in this section, the moving platform n : nGi (i = 1, 2, 3, 4) is connected to the reference platform 1 : 1Gi : 0 by four spatial complex limbs with four or five degrees of connectivity"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002278_2013-01-1491-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002278_2013-01-1491-Figure1-1.png",
        "caption": "Figure 1. Gear tilting due to radial gap between NRB and gear bore diameter",
        "texts": [
            " optimized value of lead and profile crowning and end relief (tip relief and root relief) to be designed with the consideration of all misalignments of the gearbox systems, so that the minimum TE can be achieved. One more important factor which adversely affects on the gear micro geometry and load distribution along the meshing tooth face is the gear misalignment or gear tilt due to the radial clearance between the gears bore diameter and the shaft outer diameter or outer diameter of the needle roller bearing as shown in Figure 1. Although the uniform distribution of load and center contact pattern of the meshing tooth is achieved by optimizing the gear micro geometry, by providing the crowing and tip modifications. If the radial gap between the gear and the shaft or NRB is not optimized, then the gear tilting due to this gap reduces the benefits of crowning and tip modifications and increases the gear mesh misalignment and the transmission error. This gear mesh misalignment may results in shifts in the load distribution of gear pair that results in increasing contact and bending stresses, moving the peak bending stress to the edge of the face width, and might increases gear noise"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.71-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.71-1.png",
        "caption": "Fig. 2.71 4CPaPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology C\\Pa\\\\Pa",
        "texts": [
            "22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No. 58 T ab le 2. 4 L im b to po lo gy an d co nn ec ti ng co nd it io ns of th e fu ll y- pa ra ll el so lu ti on s w it h no id le m ob il it ie s pr es en te d in F ig s. 2. 73 , 2. 74 , 2. 75 , 2. 76 , 2. 77 , 2. 78 ,2 .7 9, 2. 80 ,2 .8 1, 2. 82 ,2 .8 3, 2. 84 ,2 .8 5, 2. 86 ,2 .8 7, 2. 88 ,2 .8 9, 2. 90 ,2 .9 1, 2. 92 ,2 .9 3, 2. 94 ,2 .9 5, 2. 96 ,2 .9 7, 2. 98 ,2 .9 9, 2. 10 0, 2. 10 1, 2. 10 2, 2. 10 3, 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003670_amr.479-481.670-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003670_amr.479-481.670-Figure2-1.png",
        "caption": "Figure 2 shows the mechanical analysis of wheel hub bearing when vehicle is under driving with turning.",
        "texts": [],
        "surrounding_texts": [
            "Under the Condition of Driving with Turning. When vehicle is under driving with turning, the wheel hub bearings mainly receive the force in two directions, axial and radial. Through the analysis of load source it is clear that, the axial force mainly comes from lateral force on wheel yF , while the radial force is from three aspects, driving force on wheel DF , vertical force from ground ZF and also yF . In order to make the question more clear and simple, before we analyze the load condition of the bearing, firstly we propose the assumptions as follows: tire-road friction is good, longitudinal and lateral force don\u2019t reach the limit, don\u2019t consider the dynamic change of vertical load from road roughness, ignore effects from suspension pin, lateral force don\u2019t drift the center line of wheel, four wheels are all wheel hub motors, all motors\u2019 output torque are equal. According to the state of vehicle driving with turning, for both front and rear wheels of vehicle, ignore the load variation on both left and right wheels, and the lateral force from unsprung mass, but consider the location of gravity center moving to back, we get equations of the load on wheel [4]: 2 sin / 4 21.15 d D C A udu F m mg dt \u03b4 \u03b1  \u22c5 \u22c5 = + \u00b1    (1) ( ) ( )2 2 gr Z f r f r m du dt hmgl F l l l l = \u2212 + + (2) ( )2 s r y f r m l dv F dtl l = \u22c5 + (3) Where \u03b4 is conversion coefficient of vehicle mass, m is vehicle mass, sm is sprung mass, dC is coefficient of air resistance, u is longitudinal velocity, A is Car windward acreage, \u03b1 is road slope angle, fl is distance between gravity center and front axle, rl is distance between gravity center and rear axle, gh is height of gravity center, v is lateral velocity. external bearings. Where iF is radial force of internal bearing, oF is radial force of external bearing, e is offset distance between external bearing and wheel, l is center distance between internal and external bearings, R is wheel radius. 2 2 i D Z y e e R F F F F l l l    = + +        (4) 2 2 o D Z y l e l e R F F F F l l l \u2212 \u2212   = + \u2212        (5) Axial force of wheel hub bearing mainly comes from lateral force on wheel, which splits into two parts, axial force ' yF , in the direction of bearing axle and moment. This ' yF is equal to lateral force yF , assuming that the axial force of internal and external bearings iyF , oyF are equal. ( ) ' 4 1 2 iy oy r f r y ml dv dtl F F l F \u22c5 + = = = (6) Under the Condition of Braking with Turning. When vehicle is under braking with turning, the condition is very similar. Wheel hub bearings suffer the force in two directions, axial and radial. Axial force comes from lateral force, while radial force comes from four aspects, braking force B F , vertical force Z F , tangential force from brake caliper F\u00b5 and lateral force yF . 2 sin / 4 21.15 d B C A udu F m mg dt \u03b4 \u03b1  \u22c5 \u22c5 = \u2212    \u2213 (7) ( ) ( )2 2 gr Z f r f r m du dt hmgl F l l l l = + + + (8) ( )2 s r y f r m l dv F dtl l = \u22c5 + (9) At the same time, according to the principle of moment balance, we get the format of tangential force from brake caliper F\u00b5 , where r is effective radius of brake disc. 2 sin 21.15 dB C A uF R du R F m mg r dt r \u00b5 \u03b4 \u03b1  \u22c5 \u22c5\u22c5 = = \u2212 \u22c5    \u2213 (10) According to the condition of driving with turning we calculate the radial force of both bearings as follows, while axial force is just the same as the equation under driving with turning, where 1 l is offset distance between internal bearing and brake disc, \u03b2 is angle between brake caliper and Z axis. ( ) ( ) 2 2 1 1cos sin yB BB Z i RFl l RF l l RFeF eF F l l r l l r l \u03b2 \u03b2+ +    = + + \u2212 +   \u22c5 \u22c5    (11) 22 1 1cos sin yB B o B Z RFl RF l RFl e l e F F F l l r l l r l \u03b2 \u03b2 \u2212 \u2212 = \u2212 + + \u2212  \u22c5 \u22c5    (12)"
        ]
    },
    {
        "image_filename": "designv11_100_0002826_s1644-9665(12)60146-0-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002826_s1644-9665(12)60146-0-Figure1-1.png",
        "caption": "Fig. 1. Physical model of the Euler loaded column",
        "texts": [
            " Using a stiff cylindrical element in the load charging heads, in comparison to a rolling element, makes the rotating loop stiff between the two cooperating structure loading elements. Keywords: free vibration, stability, slender system, Euler load Conservative loads may be induced by loading or supporting structure. Depending on the construction solution we can have a jointed support, stiff fix or resilient fix of the column. The shape of the deflected axle and a physical model of the column are shown in Figure 1. This column is built from two rods with the symmetrical distribution of bend stiffness (Ei Ji) and mass per length unit ( iAi), while E1J1 = E2J2, 1A1 = 2A2 where Ei, Ji, Ai, i (i = 1, 2) are accordingly: elongated resiliency module, sectional central moment of inertia, square measure and volume weight of i\u2013 column rod. The rods are rigidly connected at both ends assuring evenness of movement and the deflection angle at the simply-supported points. Quantity c0, c2 are stiffnesses of the rotational springs at the fixing points of column",
            " The columns are labelled according to the cylindrical element in use: \u2013 column C \u2013 rolling ball bearing 5(1), \u2013 column D \u2013 stiff cylindrical element 5(2) with circular contour of the working area. A loss of stability occurs in the plane with lower bending stiffness, after exceeding a certain value of the axial force P called critical force Pc. The formulation of the vibration topic is based on the Hamilton rule using the Bernoullie-Euler theory. For this purpose we define the components of kinetic and potential energy of the elements of the structure, taking into account the physical model of the column (Figure 1), number of its rods (i = 2) and the internal forces in the rod Si (S1 = S2 = 0.5 P). The kinetic energy of the structure in Figure 1 is formulated: ., 2 1 2 0 1 i l ii i i dx t txWAT i (1) Total potential energy V1 of the column is the sum of the energy of the internal and external forces, support and fix stiffness i.e. ,3121111 VVVV (2) where : elastic strain energy ,, 2 1 2 0 2 2 11 i l i ii i i dx x txWEJV i (3) potential lengthwise energy S , , 2 1 2 0 2 1 21 j l j jj j j dx x txW SV i (4) potential energy of support and fix resiliency .. 2 1, 2 1 2 1 11 2 2 01 11 031 11 1 lx x x txWc x txWcV (5) Considering the kinetic (formula (1)) and potential energy (formulae (2\u20135)), the Hamilton rule is written as follows: "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003750_amm.197.764-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003750_amm.197.764-Figure3-1.png",
        "caption": "Fig. 3: Finite element mesh.",
        "texts": [
            " The powder feeding during cladding is realized by setting the related elements into \u201cdeath\u201d and \u201cbirth\u201d. This enhances the simulation accuracy. The dimensions of the workpiece are shown in Fig.1 and Fig.2. The material of substrate is 10 mild carbon steel (AISI1010). The clad layer is 1Cr18Ni9 stainless steel (AISI302). The thermal properties of materials are temperature-dependent [8]. The heat convection coefficient between workpiece and surrounding is 6Wm -2 K -1 . Half of the workpiece is calculated as the symmetry. The object is meshed with 8 nodes hexahedral isoparametric elements (Fig.3). The mesh around clad is refined to assure the analysis precision. 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 TTP, www.ttp.net. (ID: 129.137.5.42, University of Cincinnati, Cincinnati, USA-29/12/14,23:45:51) In laser cladding, the laser footprint moves related with the substrate and the metal powder is fed into the melten pool simultaneously. So before the laser irradiation arrives, elements at that position are not allowed be part of the model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002047_978-1-4471-5110-4_6-Figure6.20-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002047_978-1-4471-5110-4_6-Figure6.20-1.png",
        "caption": "Fig. 6.20 Problem 6.7",
        "texts": [],
        "surrounding_texts": [
            "If curl P = \u2207\u00d7P = 0 the force P is conservative. The expression \u2207\u00d7P is calculated with MATLAB: rotP_ = curl(P_,[x y z]); and the results is: curl(P_)=[0, 0, 0] The potential energy is calculated with V (x, y, z) = \u2212 \u222b P \u00b7 dr + C = \u2212 \u222b [\u2212kxdx \u2212 kydy + (k R \u2212 kz \u2212 mg)dz] + C V (x, y, z) = k x2 2 + k y2 2 + k (R \u2212 z)2 2 + mgz + C, (6.8) where C is an arbitrary constant known as the constant of integration. The equation of the sphere can be written as x2 + y2 + z2 = R2 or z = \u221a R2 \u2212 x2 \u2212 y2, (6.9) and with Eq. (6.8) the potential energy function of x and y is calculated in MATLAB with: V1=-int(P_(1)); V2=-int(P_(2)); V3=-int(P_(3)); V=V1+V2+V3+C; fVxy=subs(V,z,sqrt(R\u02c62-x\u02c62-y\u02c62)); Vxy=simple(simplify(Vxy)); and V (x, y) is obtained as V (x, y) = k R2 + (mg \u2212 k R) \u221a R2 \u2212 x2 \u2212 y2 + C. The partial derivative of the function V (x, y) with respect to x and y are \u2202V \u2202x = \u2212 (mg \u2212 k R) x\u221a R2 \u2212 x2 \u2212 y2 , \u2202V \u2202y = \u2212 (mg \u2212 k R) y\u221a R2 \u2212 x2 \u2212 y2 . The equilibrium positions of the particle are obtained from \u2202V \u2202x = 0, \u2202V \u2202y = 0. In MATLAB the equilibrium positions are obtained with: dVxydx = simple(diff(Vxy,x)); dVxydy = simple(diff(Vxy,y)); xe=solve(dVxydx,x); ye=solve(dVxydy,y); ze=solve(xe\u02c62+ye\u02c62+z\u02c62-R\u02c62,z); The results for the equilibrium positions are M1(0, 0, R) and M2(0, 0,\u2212R). Example 6.5 A particle P of mass m is on a circle of radius R as shown in the Fig. 6.13. The circle is on a vertical plane xy. Find the equilibrium positions of the particle. Solution The independent variable is the angle \u03b8. The position of the particle P is x = R*cos(theta); y = yN+R+R*sin(theta); r_ = [x y]; where yN is the y coordinate of the lower end N of the circle. The gravity is the only force acting on the particle and the potential energy is calculated with: dr_=diff(r_,theta); G_ = [0 -m*g]; V = -int(G_*dr_.\u2019); fprintf(\u2019V=%s + C\\n\u2019, char(V)) The MATLAB expression for the potential energy is: V=R*g*m*sin(theta) + C where C is a constant of integration. The equilibrium positions are calculated from the equation: dV = diff(V,theta); thetae=solve(dV,theta); theta1=thetae; theta2=theta1+pi; The equilibrium position are the points M and N as shown in Fig. 6.13: theta1 = pi/2 and theta2 = (3 \u2217 pi)/2. The equilibrium stability is verified with the second derivative of the potential energy: d2V = diff(dV,theta); d2V1=subs(d2V,theta,theta1); d2V2=subs(d2V,theta,theta2); and the MATLAB results are: d2V/d(theta)\u02c62=-R*g*m*sin(theta) for theta1 => d2V/d(theta)\u02c62=-R*g*m for theta2 => d2V/d(theta)\u02c62=R*g*m The equilibrium position \u03b8 = 3 \u03c0/2, position N , is a stable equilibrium because d2V/d\u03b82 = Rgm is positive. 6.6 Problems B A C 1 2 k \u03b8 m, l m, l F A B A C \u03b8 m, l l k k Fig. 6.18 Problem 6.5 B A k O C F \u03b8 Fig. 6.19 Problem 6.6 B A k \u03b8 m, l k Fig. 6.21 Problem 6.8 \u03b8 l m, l m, l m, l F B A C D k 1 2 3 0 0 by the torsion spring is M = K \u03b8, where \u03b8 is the relative angle between the links at the joint. Determine the minimum value of K which will ensure the stability of the mechanism for \u03b8 = 0. 6.8 Figure 6.21 shows a four-bar mechanism with AD = l. Each of the links has the mass m (m1 = m2 = m3 = m) and the lenght l (l1 = l2 = l3 = l). At B a vertical force F acts on the mechanism and the spring stiffness is k. The motion is in the vertical plane. Find the equilibrium angle \u03b8. Use the following numerical application: l = 15 in, m = 10 lb, F = 90 lb, and k = 15 lb/in. Select an unextended (initial) length L0 for the spring. 6.9 A particle of mass m can move freely in space. The potential energy V of the particle at x = l, when the particle is subject to a vertical force F = ax2 + bx + c, is V = s. Find the equilibrium positions of the particle. For the numerical application use a = 1, b = \u22123, c = 0, l = 0 m, and s = 5 J. 6.10 A bar of mass m and length l is supported by a vertical wall and a point at O , as shown in Fig. 6.22. Find the equilibrium positions of the bar. For the numerical application use l = 0.5 m, a = 0.1 m, m = 1 kg, and g = 9.81 m/s2."
        ]
    },
    {
        "image_filename": "designv11_100_0002152_978-3-642-40852-6_39-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002152_978-3-642-40852-6_39-Figure4-1.png",
        "caption": "Fig. 4. The hexapod (6-UPS) parallel manipulator",
        "texts": [
            " Based on the modified updating theme, the evolving population will search the best position for the particles within the feasible region all the time to guarantee the algorithm proceed successfully. Additionally, the restitution coefficient can be modified flexibly between 0 and 1 in different cases. As pointed by Nategh [14], the maximum observability index is potentially obtained at the extreme boundary of the robots\u2019 motion constraints. Then, for our optimal configurations selection problem, the collisions are supposed to be perfectly inelastic ones, namely CR,j = 0, to make the algorithm quickly converge. In this section, the hexapod (6-UPS) parallel manipulator, as shown in Fig.4, is studied as a numerical example to demonstrate the proposed calibration method. There are 7 independent kinematic parameters to be identified in each limb, namely the position vector of the universal joint on the fixed platform u i, the initial length of the prismatic joint li and the position of the spherical joint on the moving platform si. Thus, there are totally 42 parameters to be identified during the kinematic calibration, whose nominal values are listed in Tab.1. Each measurement generates six constraint equations to the calibration system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002691_imece2012-87321-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002691_imece2012-87321-Figure3-1.png",
        "caption": "Figure 3. NON-TORQUE GEARBOX LOADING",
        "texts": [
            " The input for each of these FRFS are the forces at the rotor main shaft coupling. Figure 2 shows the rotor-main shaft coupling and the corresponding forces. The main response DOFs are the planet carrier displacements at the rear planet carrier (PLC) bearing location (PLC-B) and the gearbox bushing displacements. Figure 1 shows the different bearings in the gearbox, whereas Figure 2 shows the gearbox bushings. The planet carrier motion is investigated, since this parameter is significantly affected by the non-symmetric loading of the gearbox. Figure 3 illustrates this. Due to the planet carrier displacements the loading conditions in the planet-ring gear meshes can become unfavorable. This can potentially result in non-symmetric planet bearing loading and corresponding overloading of one of the planet bearings [2]. The planet-ring gear meshes displacements are not used in the FRF analysis in this paper since we want to make abstraction of the influence of the gear meshing stiffness on the behavior of the gearbox suspension. In essence the gearbox should be designed such that the non-torque loading is transferred from the planet carrier through the planet carrier bearings to the housing"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.12-1.png",
        "caption": "Fig. 7.12 A 2D torus T 2 situated in Euclidean spaces R 3 and R 2",
        "texts": [
            " Finally, the map f3 : S1 \u2192 R 3, as shown in Figure 7.11, is always an embedding, because in this case, n = 1 and m = 3 = 2n+1 so that f3 meets the condition of the Whitney Theorem. The second example is a regular 2-torus T 2. The surface equation represented in R 3 is given by \u23a7\u23a8 \u23a9 z1 = (a+ r cos \u03b82) cos \u03b81 z2 = (a+ r cos \u03b82) sin \u03b81 z3 = r sin \u03b82, (7.41) where a > 0 and r > 0 are two constant parameters, called the major and minor radii, respectively, and the two local coordinates q1 = \u03b81 and q2 = \u03b82, as shown on the left of Figure 7.12. It can be easily observed that if a = 0, equation (7.41) represents a 2-dimensional spherical surface S2 of radius r, while if a > r, this is a smooth and compact surface of 2-torus T 2. Topologically, this torus is viewed as a product of two 1-circles that are corresponding to the two rotation angles \u03b81 and \u03b82, i.e., T 2 S1\u00d7S1 [27, 28]. The Jacobian matrix J of (7.41) becomes J = \u2202\u03b6 \u2202q = \u239b \u239d \u2212(a+ r cos \u03b82) sin \u03b81 \u2212r sin \u03b82 cos \u03b81 (a+ r cos \u03b82) cos \u03b81 \u2212r sin \u03b82 sin \u03b81 0 r cos \u03b82 \u239e \u23a0 . (7.42) Thus, the Riemannian metric on the 2-torus is a diagonal matrix, W = JTJ = ( (a+ r cos \u03b82) 2 0 0 r2 ) (7.43) which is always non-singular as a > r > 0. Therefore, the mapping \u03b6 : T 2 \u2192 R 3 given by (7.41) is an embedding. In fact, it can be shown [27] that all the known compact and orientable 2- dimensional manifolds can be embedded into R 3. However, if T 2 is mapped into a Euclidean 2-space R2, which is, for instance, spanned only by z1 and z2 axes, as shown on the right of Figure 7.12, the explicit form of this mapping \u03b62d is just the first two equations of (7.41), and its Jacobian matrix takes the first two rows of (7.42), as if the third equation of z3 in (7.41) and the third row for the Jacobian matrix in (7.42) disappear. Then, the determinant of this 2 by 2 Jacobian matrix is r(a+ r cos \u03b82) sin \u03b82, which vanishes at \u03b82 = 0 or \u03c0. This evidently shows that on the 2D image of the torus T 2, each point on either the inner or the outer circular contour is a singular point. Therefore, the mapping \u03b62d : T 2 \u2192 R 2 is neither an immersion, nor an embedding",
            " The multi-configuration issue will be very important for our next discussion on robotic dynamics and adaptive control, and the above example is just an explicit interpretation at both the geometrical and topological standpoints. Unfortunately, for an RR-type 2-link planar arm, or a robot having two revolute joints, whose axes are parallel to each other, its C-manifold becomes a flatted torus, as shown in Figure 7.13. This is also similar to the above case of squeezing T 2 into the 2D plane, resulting in an annulus in Figure 7.12. The z3 component of the original T 2 is now pushed down into the 2D plane to impose an additional projection on each of the first two components z1 and z2. According to Figure 7.13 on the right at point P , those additional projections are \u2212z3 sin \u03b81 = \u2212r sin \u03b82 sin \u03b81 for z1, and z3 cos \u03b81 = r sin \u03b82 cos \u03b81 for z2 such that the new equation for the flatted torus turns out to be { z1 = (a+ r cos \u03b82) cos \u03b81 \u2212 r sin \u03b82 sin \u03b81 = a cos \u03b81 + r cos(\u03b81 + \u03b82) z2 = (a+ r cos \u03b82) sin \u03b81 + r sin \u03b82 cos \u03b81 = a sin \u03b81 + r sin(\u03b81 + \u03b82)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002393_roman.2013.6628492-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002393_roman.2013.6628492-Figure1-1.png",
        "caption": "Figure 1. Configuration of pipeline type sterilizer.",
        "texts": [
            " Waterway type UV sterilizer is installed in open channel. It is suitable for big capacity of sewage treatment. Pipeline type UV sterilizer is inserted into the middle of a conduit. It is applied to a wide variety of fields for many years. It also has been in the limelight for reasons of maintenance convenience, economical price, and safety. In particular, it is applied enlarged in sewage treatment unit to small village. In a pipeline type UV sterilizer, disinfecting UV lamps are primarily located in the conduit as shown in Figure 1. It is difficult to know the lamp operation from the outside. Therefore, it is necessary to set up an external control panel to monitor the situation in the pipeline, such as for maintenance or operation of the lamp. However, It is not only necessary the cost to install of these panels but also necessary another space to set up. Besides, additional control panel is needed when *This work (Grants No C0027764.) was supported by Business for Cooperative R&D between Industry, Academy, and Research Institute funded Korea Small and Medium Business Administration in 2012 ",
            " Jaebyung Park is with the Chonbuk National University, Jeonju, ASI | KR | KS004 | JEONJU (corresponding author, phone: +82-63-270-4283; fax: +82-63-270-2394; e-mail: jbpark@ jbnu.ac.kr) . adding a pipeline type UV sterilizer. Thus, it is not efficient in many aspects. In this paper, smart device interface for intelligent control of pipeline type UV sterilizer is proposed. It is more economical and convenient than the existing control panel by using smart device to monitor and control of a pipeline type UV sterilizer. The configuration of pipeline type sterilizer is shown in Figure 1. UV lamps for disinfection are located in the pipeline and a lamp guard is covering the lamp. The lamp guard is made of quartz. It looks like a glass tube. The wiper cleans the lamp guard when scraps cover the lamp guard and reduce the sterilizing power. The lamp cleaning wiper moves forward and backward in horizontal axis by rolling drive screw. There is a motor and gearbox to roll the drive screw. UV power sensor observes sterilizing power by checking UV power. There is a relationship between UV dose amount and sterilizing power"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.5-1.png",
        "caption": "Figure 12.5 A turning disc about its diagonal on a turning table.",
        "texts": [
            " The disc makes the mass moment of the table asymmetric: 0I1 = 0I2 = 0I3 (12.126) Substituting the constant angular velocity 0\u03c91 from (12.112) into Euler equations (12.10)\u2013(12.12), we have 0M1 = 0 0M2 = 0 0M3 = 0 (12.127) So, as long as is constant, no moment is needed to keep the table turning. z0 \u03b82 \u03b81 z2x1 B1 B2 l z1 x2 y1 R y2 F F Figure 12.4 Bearing forces F on the shaft. Example 722 A Turning Disc about a Diagonal on a Turntable Let us change the direction of the disc of Figure 12.3 to be mounted as shown in Figure 12.5. The uniform disc has a mass m, radius R, and mass moment [I ]. The disc is turning about its diagonal with constant angular velocity \u03b8\u03072 = \u03c9 in B1, which is a fixed coordinate frame on the table at the mass center of disc C. The table is also turning with angular velocity \u03b8\u03071 = in the global frame B0. The mass center of the disc is motionless. We attach a principal coordinate frame B2 to the disc and express Euler equations in B2. The transformation matrices between the coordinate frames are given in Equations (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure1-1.png",
        "caption": "Fig. 1. Drafting View of Design 1 with Dimensions",
        "texts": [],
        "surrounding_texts": [
            "The detailed CAD designs of the PAMD using the NX 7.5 software are shown in Figures 1 and 2 with dimensions."
        ]
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure28-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure28-1.png",
        "caption": "Fig. 28 Heel chamfering with BM\u2014spiral bevel pinion",
        "texts": [
            " Beyond this, even a small diameter EndMill is likely to interfere with the concave tooth flank when chamfering the bottom of the tooth (right, Fig. 26) and therefore either the Pivot Angle must be reduced, or else this solution becomes unacceptable. Tool spindle to turn table collision is not likely to occur at Heel (left, Fig. 27); and if the Pivot Angle is correctly chosen, tool interference with the tooth flank can be avoided (right, Fig. 27). The Ball Mill tool (BM), thanks to its spherical end, can be fitted in places where an End Mill tool would not do an acceptable job. Consider for example the spiral bevel pinion shown in Fig. 28, left. The Ball Mill tool can be plunged vertically along the Toe and Heel edges without any risk of tool spindle to turn table interference. And by carefully selecting the Ball Mill diameter, the fillet area can also be chamfered. Five unit vectors (left, Fig. 29) are required to control the BM at any point along a tooth edge: Vo: N X T ,\u2212\u2212\u2192 Tool: the tool vector,\u2212\u2212\u2212\u2192 Trans: axis about which vector \u2212\u2212\u2192 Tool is rotated by the Pivot Angle. \u2212\u2192 Vo is obtained from the cross product of N and T . Again, vector \u2212\u2212\u2212\u2192 Trans is obtained by pivoting T about \u2212\u2192 Vo by \u03c0 2 + the local pressure angle; vector \u2212\u2212\u2192 Tool is obtained from the cross product of \u2212\u2212\u2212\u2192 Trans and \u2212\u2192 Vo; vector \u2212\u2212\u2192 Tool can be pivoted parallel to itself about the local tooth edge point by the Chamfer Angle\u2014left, Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003995_mechatron.2011.5961067-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003995_mechatron.2011.5961067-Figure1-1.png",
        "caption": "Fig. 1. Differentiation of a conductor into layers.",
        "texts": [
            " in [12] and it divides the whole conductor into many parallel layers, calculates electric properties of every layer (the inductance and the resistivity of the layer) and solves an equivalent circuit of winding bar. The advantage of this method is the fact that the calculation is not based on solution of magnetic field (like common calculations considering eddy currents), but solves an electric problem and as next step the magnetic field of the conductor may be obtained as result of this calculation. The differentiation of a trapezoid shaped slot is shown in Fig. 1. The whole conductor is divided into parallel layers with equal heigh dh and variable width bi. These parallel layers are considered as rectangles and the resistance of each rectangle is r i= 1 \u03b3 Al \u22c5 l bidh , where \u03b3Al is conductance of bar's material l is length of the bar bi is width of the layer and dh is its height. (1) Similarly to (1) the inductance of one layer is \u03bb i=\u03bc0 l dh bi , where \u03bc0 is permeability of the vacuum. Considering current frequency in the bar the resulting leakage reactance of each layer the bar is xi=\u03c9\u03bb i , where \u03c9 is angular frequency of the current"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003400_iccve.2013.6799905-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003400_iccve.2013.6799905-Figure4-1.png",
        "caption": "Fig. 4. Frame Stress Analysis",
        "texts": [],
        "surrounding_texts": [
            "The detailed CAD designs of the PAMD using the NX 7.5 software are shown in Figures 1 and 2 with dimensions."
        ]
    },
    {
        "image_filename": "designv11_100_0001202_tec.2021.3058804-Figure9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001202_tec.2021.3058804-Figure9-1.png",
        "caption": "Fig. 9. Flux lines and current distribution for given iqds and iqdseg by varying \u03b8e and \u03c7.",
        "texts": [],
        "surrounding_texts": [
            "PERIODICITY IN INDUCTANCE MATRICES AND FLUX Lemma 2. The segment qd inductance matrices are periodic with respect to \u03b8e, with period \u03c0/3 radians: Lqdseg,seg(iqds, iqdseg, \u03b8e + \u03c0 3 , \u03c7) = Lqdseg,seg(iqds, iqdseg, \u03b8e, \u03c7) , (40) Lqdseg,s(iqds, iqdseg, \u03b8e + \u03c0 3 , \u03c7) = Lqdseg,s(iqds, iqdseg, \u03b8e, \u03c7) . (41) Proof. The segment fluxes are expressed as (see (9) of Part I), \u03bbseg = Lseg,seg(is, iseg, \u03b8r) iseg + Lseg,s(is, iseg, \u03b8r) is , (42) where Lseg,seg and Lseg,s are segment-to-segment and segmentto-stator inductance matrices, respectively. Equivalently, we write \u03bbseg as a function of iqds, iqdseg, \u03b8e, and \u03c7, i.e., \u03bbseg = Lseg,seg(iqds, iqdseg, \u03b8e, \u03c7)T \u22a4 r (\u03c7) iqdseg + Lseg,s(iqds, iqdseg, \u03b8e, \u03c7)T \u22a4 s (\u03b8e) iqds . (43) Keeping iqds, iqdseg, and \u03c7 fixed to arbitrary values, the field distribution at \u03b8e and \u03b8e+\u03c0/3 is illustrated in Figs. 9a and 9b, respectively. It can be observed that the rotor experiences exactly the same field distribution at these two positions, i.e., \u03bbseg(iqds, iqdseg, \u03b8e + \u03c0 3 , \u03c7) = \u03bbseg(iqds, iqdseg, \u03b8e, \u03c7) . (44) This always holds when the number of stator slots per pole per phase is an integer. Using (43) in (44), and hiding the dependence of the inductances on iqds and iqdseg for notational brevity, yields Lseg,seg(\u03b8e + \u03c0 3 , \u03c7)T \u22a4 r (\u03c7) iqdseg + Lseg,s(\u03b8e + \u03c0 3 , \u03c7)T \u22a4 s (\u03b8e + \u03c0 3 ) iqds = Lseg,seg(\u03b8e, \u03c7)T \u22a4 r (\u03c7) iqdseg + Lseg,s(\u03b8e, \u03c7)T \u22a4 s (\u03b8e) iqds . (45) Multiplying the above equation with Tr(\u03c7) from the left on both sides, and using the qd inductance matrix definitions Lqdseg,seg = TrLseg,segT \u22a4 r , Lqdseg,s = TrLseg,sT \u22a4 s , (46) yields Lqdseg,seg(\u03b8e + \u03c0 3 , \u03c7) iqdseg + Lqdseg,s(\u03b8e + \u03c0 3 , \u03c7) iqds = Lqdseg,seg(\u03b8e, \u03c7) iqdseg + Lqdseg,s(\u03b8e, \u03c7) iqds. (47) For the above equality to hold for arbitrary iqds and iqdseg, it is necessary that Lqdseg,seg(\u03b8e + \u03c0 3 , \u03c7) = Lqdseg,seg(\u03b8e, \u03c7) , (48) Lqdseg,s(\u03b8e + \u03c0 3 , \u03c7) = Lqdseg,s(\u03b8e, \u03c7). Lemma 3. The segment qd inductance matrices are periodic Authorized licensed use limited to: London School of Economics & Political Science. Downloaded on May 16,2021 at 23:17:07 UTC from IEEE Xplore. Restrictions apply. 0885-8969 (c) 2021 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. with respect to slip angle \u03c7, with period 2\u03c0/n radians: Lqdseg,seg(iqds, iqdseg, \u03b8e, \u03c7+ 2\u03c0 n ) = Lqdseg,seg(iqds, iqdseg, \u03b8e, \u03c7) , (49) Lqdseg,s(iqds, iqdseg, \u03b8e, \u03c7+ 2\u03c0 n ) = Lqdseg,s(iqds, iqdseg, \u03b8e, \u03c7) . (50) Proof. Keeping iqds, iqdseg, and \u03b8e fixed to arbitrary values, the field distribution at \u03c7 and \u03c7+2\u03c0/n is illustrated in Figs. 9a and 9c, respectively. The rotor has turned by one rotor slot pitch in the clockwise direction; however, the stator experiences exactly the same field distribution in both cases. Therefore, \u03bbseg(iqds, iqdseg, \u03b8e, \u03c7+ 2\u03c0 n ) = Ur\u03bbseg(iqds, iqdseg, \u03b8e, \u03c7) , (51) where Ur is a square block diagonal matrix, Ur = blockdiag(P3,P3, \u00b7 \u00b7 \u00b7 ,P3) , (52) and P3 is permutation matrix of size n\u00d7 n given by P3 =   0 0 \u00b7 \u00b7 \u00b7 0 1 1 0 \u00b7 \u00b7 \u00b7 0 0 0 1 \u00b7 \u00b7 \u00b7 0 0 ... ... . . . ... ... 0 0 \u00b7 \u00b7 \u00b7 1 0   . (53) It can be readily shown using trigonometric identities that Ur = T\u22a4 r (\u03c7+ 2\u03c0 n )Tr(\u03c7) . (54) Using (43) in (51), and hiding the dependence of inductances on iqds and iqdseg for notational brevity, we have Lseg,seg(\u03b8e, \u03c7+ 2\u03c0 n )T\u22a4 r (\u03c7+ 2\u03c0 n ) iqdseg + Lseg,s(\u03b8e, \u03c7+ 2\u03c0 n )T\u22a4 s (\u03b8e) iqds = T\u22a4 r (\u03c7+ 2\u03c0 n )Tr(\u03c7)Lseg,seg(\u03b8e, \u03c7)T \u22a4 r (\u03c7) iqdseg +T\u22a4 r (\u03c7+ 2\u03c0 n )Tr(\u03c7)Lseg,s(\u03b8e, \u03c7)T \u22a4 s (\u03b8e) iqds . (55) Multiplying the above equation with Tr(\u03c7+ 2\u03c0 n ) from the left on both sides, and using the definitions (46), we obtain Lqdseg,seg(\u03b8e, \u03c7+ 2\u03c0 n ) iqdseg + Lqdseg,s(\u03b8e, \u03c7+ 2\u03c0 n ) iqds = Lqdseg,seg(\u03b8e, \u03c7) iqdseg + Lqdseg,s(\u03b8e, \u03c7) iqds . (56) For the above equality to hold for arbitrary iqds and iqdseg, it is necessary that Lqdseg,seg(\u03b8e, \u03c7+ 2\u03c0 n ) = Lqdseg,seg(\u03b8e, \u03c7) , (57) Lqdseg,s(\u03b8e, \u03c7+ 2\u03c0 n ) = Lqdseg,s(\u03b8e, \u03c7) . Corollary 1. The reduced qd inductance matrices L\u2212 qdr,seg and L\u2212 qdr,s are periodic with respect to \u03b8e and slip angle \u03c7, with period \u03c0/3 and 2\u03c0/n radians, respectively; for example, L\u2212 qdr,seg(iqds, iqdseg, \u03b8e + \u03c0 3 , \u03c7) = L\u2212 qdr,seg(iqds, iqdseg, \u03b8e, \u03c7) . (58) Proof. The reduced inductance matrices are obtained by linear transformations of Lqdr,seg and Lqdr,s with a constant matrix C (see Section III-C of Part I). These are related to segment inductance matrices by (17)\u2013(18), where Le, P\u03031, P2, M\u0303\u22121, and N\u0303 are constant. The corollary is a direct consequence of Lemmas 1 and 2. Corollary 2. The flux ripple \u03bb\u0303\u2212 qdr0 and the incremental inductance matrix Linc\u2212 qdr are periodic with respect to \u03b8e and slip angle \u03c7, with period \u03c0/3 and 2\u03c0/n radians, respectively. Proof. The reduced segment flux \u03bb\u2212 qdr (see (48) of Part I) is periodic because of Corollary 1. Hence, \u03bb\u0303\u2212 qdr0 and Linc\u2212 qdr,seg are periodic."
        ]
    },
    {
        "image_filename": "designv11_100_0003419_amr.199-200.449-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003419_amr.199-200.449-Figure2-1.png",
        "caption": "Fig. 2 The configuration of Hall elements in the static oblique coordinates",
        "texts": [
            " The main objective of this paper is to propose a new and simple method, namely the static oblique coordinates method, to measure bearing balls\u2019 3D motion from static measure position based on the work of Akira Nakashima\uff0cZhang[12]et al. Another objective is to propose a method to uniformly magnetize a steel ball and detect the magnetic pole and magnetism of the magnetized ball. In the static oblique coordinates method\uff0cthe magnetized balls are assembled in the test bearing. The Hall sensors are installed on the outer race and the side of cage (fig. 2). These devices are used to measure the 3D motion of steel balls in an angular contact ball bearing. Principles of Measurement. The basic principle of this measure method is given as follows. First, some magnetized balls are assembled into a ball bearing. Then, a magnetic circuit is composed by the magnetized balls together with the Hall sensors which are installed on the outer race and the side of cage (fig. 2). The voltages outputted from the Hall sensors are proportional to the magnetic field intensity. When the magnetic balls roll in the bearing, the direction of magnetic poles varied, causing the variation of the Hall sensors\u2019 output voltages. Thus, the balls\u2019 3D motion can be derived from the measurement of the variation of the magnetized balls\u2019 magnetic poles in three directions of coordinate axis. Figure 2 shows the static oblique coordinates. The ball\u2019s center is set as the original point. The directions of three static coordinate axis X, Y\u2019, Z are showed in fig. 2. Figure 3 shows the schematic diagram of the magnetic pole measure principle using Hall sensors. The voltages of the Hall sensors in the three directions of coordinate axis VHx, VHy, and VHz are given as cosHx oV V \u03b1= \uff081\uff09 cosHy oV V \u03b2= \uff082\uff09 cosHz oV V \u03b3= \uff083\uff09 where \u03b1, \u03b2, and \u03b3 are included angles between the north magnetic pole and x, y and z axis respectively, and Vo is the maximum Hall voltage. The position of the magnetized ball\u2019s north magnetic pole is )coscoscos( \u03b3\u03b2\u03b1 kjirN ++= \uff084\uff09 where r is the ball\u2019s radius; i, j and k are the unit direction vectors"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003644_amr.605-607.1176-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003644_amr.605-607.1176-Figure1-1.png",
        "caption": "Fig. 1. Model of eccentric shaft",
        "texts": [],
        "surrounding_texts": [
            "Using FEM method to make modal analysis of structure which has successive mass,the structure should be discreted into finite element. Find out the element stiffness matrix K and element mass matrix M. The node motion equation is as follow: )( 2 2 tPK dt d C dt d M =++ \u03b4 \u03b4\u03b4 . (1) If dt d dt d \u03b4 \u03b4 \u03b4 \u03b4 == , 2 2 the (1) reduces to: )(tPKCM =++ \u03b4\u03b4\u03b4 . (2) \u03b4 is acceleration vector;\u03b4 is velocity vector;\u03b4 is displacement vector, )(tP is load matrix;C is damping matrix. In (2), making P(t)=0,get free vibration equation. The influence of damping is so tiny, it shouldn't really deflect them very much,and get non damping free vibration equation: 0=+ \u03b4\u03b4 KM . (3) If the struct play simple harmonic motion: t\u03c9\u03d5\u03b4 cos= . (4) Take(4)into(3),and get homogeneous equation: 0)( 2 =\u2212 \u03d5\u03c9 MK . (5) So get natural frequency of vibration: 0 2 =\u2212 MK \u03c9 . (6) Both mass matrix and stiffness matrix are N order square,and N is the number of DOF. So the natural frequency are given. n\u03c9\u03c9\u03c9\u03c9 \u2264\u2026\u2264\u2264\u2264 321 In the condition of mass-normalized, the amplitude of a group of nodes are obtained by (5) for every natural frequency of vibration: [ ]Tiniiii \u03c6\u03c6\u03c6\u03c6\u03d5 \u2026= 321 And get vector that`s the mode of vibration[2]."
        ]
    },
    {
        "image_filename": "designv11_100_0003224_amr.565.171-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003224_amr.565.171-Figure2-1.png",
        "caption": "Fig. 2 The temperature distribution of the oil film in different rotation speed",
        "texts": [
            " In this case, the study on dynamic pressure effect of Liquid hybrid bearing in different rotational speeds is very important to analyze dynamic load capability, dynamic stability and temperature rising of the bearing. In this paper, keeping eccentric ratio 0.5\u03b5 = , the supply pressure s 7 MPaP = and supply oil temperature o25 CT = unchanged, analysis results of thermal characteristics in different rotational speeds were obtained through changing the rotational speed of spindle. The temperature field distributions of the oil film are shown as Fig.2, and the law of temperature change is shown as Fig.3. Fig. 2 shows that the highest temperature region appears in eccentric position (Minimum oil film thickness), temperature gradually increasing from the middle to both sides, and with the rotation speed increasing, the temperature generally increases and area of high temperature region gradually expands. Fig. 3 shows that the influence of rotation speed on oil film temperature rising is obviously, with the rotation speed increasing, the temperature of oil film increases constantly, and the maximum temperature changes more obviously"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002417_amr.201-203.818-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002417_amr.201-203.818-Figure3-1.png",
        "caption": "Fig. 3 Computational domain 3D solid model Fig. 4 Computational domain mesh dissection",
        "texts": [
            "Finally, the high temperature air is pushed out the motor by the centrifugal cutter. Radial ventilation duct physical model mainly includes the fluid inside the stator ventilation duct, the fluid inside the rotor ventilation duct, the fluid inside the rotor spider,ventilation channel,stator winding and rotor copper bar 6 parts,is shown in Fig. 2. Select the radial ventilation duct which between the first two iron layers close to the CDE drive end for research object.3D solid model of the flow field inside the ventilation duct is shown in Fig. 3;the mesh grid is shown in Fig. 4. Divide the solved region into about 2.36 million finite volume,0.51 million node. Conservation equations of mass and momentum. In view of the steady-state flow, the conservation equation of mass and momentum in the relative reference coordinate system can be expressed respectively, as follows: ( u ) 0r\u03c1\u2207 = \uff081\uff09 ( ) (2 ) p F\u03c1 \u03c1 \u03c4\u2207 + \u00d7 + \u00d7 \u00d7 = \u2212\u2207 +\u2207 + r r r u u \u2126 u \u2126 \u2126 r \uff082\uff09 Where \u03c1 is density; r u is the relative velocity vector, \u2207 is divergence,namely ( ) ( )div\u03c1 \u03c1\u2207 = r r u u ; \u2126 is the rotational angular velocity vector, r is the position vector of elemental volume in the rotating reference frame; p is static pressure on the air elemental volume"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.29-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.29-1.png",
        "caption": "Figure 10.29 Phasor diagram of PM synchronous motors: (a) SPM; (b) IPM; and (c) flux weakening mode of IPM",
        "texts": [
            " When the stator field is leading the rotor field, the stator will attract the rotor magnets. The machine then operates as a motor. When the stator field is lagging the rotor field, the machine becomes a generator. At no load, the rotor magnetic field will generate a back emf Eo in the stator windings. When a voltage with the same frequency is applied to the stator windings, then a current will be generated and the voltage equation can be written as V = Eo + IR + jIX (10.68) where R is the stator resistance and X is the synchronous impedance. The phasor diagram is shown in Figure 10.29 when neglecting the stator resistance. From the diagram, the term jIX can be further decomposed into two components: jIdXd and jIqxq . In fact, in IPM motors, the d axis and q axis will have different reactances. By using Figure 10.29, Equation 10.68 can be rewritten for IPM motors as V = Eo + IR + jIdXd + jIqXq (10.69) The real power can be calculated, since from Figure 10.29, \u03d5 = \u03b4 + \u03b8 : P1 = mIV cos \u03d5 = mIE o cos \u03b4 = mV (I cos \u03b4 cos \u03b8 \u2212 I sin \u03b4 sin \u03b8) = mV (Iq cos \u03b8 \u2212 Id sin \u03b8) (10.70) where \u03d5 is the power factor angle (the angle between the voltage and current), \u03b8 is the angle between the voltage and back emf, and \u03b4 is the inner power angle (the angle between the back emf and voltage). From Figure 10.29, IqXq = V sin \u03b8 IdXd = V cos \u03b8 \u2212 Eo (10.71) Therefore, the power of PM motors can be expressed as P = mEoV Xd sin \u03b8 + mV 2 2 ( 1 Xq \u2212 1 Xd ) sin(2\u03b8) (10.72) The torque can be derived by dividing Equation 10.72 by the rotor speed as shown in Figure 10.30, where the torque\u2013speed characteristics of a typical PM motor are shown. For SPM motors, since Xd = Xq , the second term of Equation 10.72 is zero. For IPM motors, the q axis has less reluctance due to the existence of soft iron in its path, and the d axis has magnets in its path which has larger reluctance",
            "74 and maintain maximum torque output at the same time. This operation is also called constant torque operation. It can also be seen from Equation 10.72 that for a given \u03b8 , the power is inversely proportional to frequency, since V , Xd , and Eo are all proportional to frequency \u03c9. This is similar to the V /f control of induction motors. When stator voltage reaches its maximum, Equation 10.74 can no longer be maintained. As \u03c9 increases, V becomes constant, and a current in the d-axis direction must be supplied, as shown in Figure 10.29c. The relationship of voltage and frequency can be expressed as V 2 = (Eo \u2212 IdXd) 2 + (IqXq) 2 = (k\u03c9\u03d5 \u2212 Id\u03c9Ld) 2 + (Iq\u03c9Lq) 2 V \u03c9 = \u221a (k\u03d5 \u2212 IdLd)2 + (IqLq)2 (10.75) This operation is also called the flux weakening operation region because the d-axis current generate a magnetic flux in the opposite direction to the PM field. Note that due to constraints such as the current limit of the inverter, the q-axis current may have to be decreased from its rated value so that the total current from the inverter is kept the same",
            " The proposed methods are validated by FEA and experiments [91]. In the following, the formulas will be derived based on a set of assumptions and then modified based on practical design considerations. The assumptions include the following: \u2022 Magnetic pole salience can be neglected \u2022 The stator resistance is negligible. \u2022 Saturation can be neglected. \u2022 The air-gap flux is sinusoidally distributed. Based on the above assumptions, and using the phasor diagram of a PM synchronous motor as shown in Figure 10.29, the input power of the PM synchronous motor can be written as P1 = mIV cos \u03c6 = mIE 0 cos \u03b4 (10.95) \u2217 \u00a9 [2006] IEEE. Reprinted, with permission, from IEEE on Magnetics. where m is the number of phases, I and V are the phase voltage and phase current, Eo is the induced back emf per phase, \u03d5 is the power angle, that is, the angle between phasor I and phasor V , and \u03b4 is the inner power angle, that is, the angle between phasor I and phasor Eo. The back emf of a PM synchronous machine with sinusoidal air-gap flux can be expressed as E0 = \u221a 2\u03c0Kwf W (10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002984_ievc.2012.6183234-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002984_ievc.2012.6183234-Figure3-1.png",
        "caption": "Figure 3. Kart Control System Hardware",
        "texts": [
            " The student was able to edit the original Arduino kart control program, and include his algorithm into it. The student also learned how to control an LCD screen to output the speed in MPH, and KPH. All of this was completed in about half the time it took to implement the LabVIEW wheel speed sensor. Based on these findings, and the findings from other seniors who were also taught LabVIEW, the final decision to use the Arduino was made. A base curriculum for an Arduino kart control system class was created (Fig. 3). The new electric kart class began in the Since the program is still in its infancy, there and improvements that are under consideration. program is a 10 week 25 total hour lab-class. students begin by learning the basics of the to program various inputs and outputs. They manipulate the outputs, and then the inputs. combined to create simple control systems. The combining the external sensors and circuits desired results. The end goal for the students is to create using the given inputs. The control system will manipulating the kart motor controller"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003430_05698195808972339-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003430_05698195808972339-Figure4-1.png",
        "caption": "FIG. 4. Hertz stress ellipsoid.",
        "texts": [
            " The rolling coefficient of friction for hard steel balls on hard steel plates is of the order of 5 X 10-5 and the energy released due to this is only 1-5 %of the energy released due to spinning. Besides the rolling friction is probably due mostly to hysteresis and hence the energy ~ ~ is not released at the surface but is ~ , . spread out in a significant volume of ,, ~ , --- '~,_ metal. This rolling energy is not / /, --- .... . considered in the following develop/ / / / .... --- --- 7 ment. Fig. 4 shows the Hertz stress /:--- .... / / / / ....- --- --- / ellipsoid at the contact zone. It will ,f::.... - - --- - - . - . --i-- . --. -?- - - - - - - . - - - . - - -1-- .-I be noted that with spinning about the (A) (B) (C) center of the contact zone the velocity FIG. 2. Showing the impossibility of pure rolling in a thrust ball of sliding of a point within the zone is zero at the center bearing. and a maximum at the periphery; while the stress is zero at the periphery and a maximum at the center"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003771_gcis.2012.37-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003771_gcis.2012.37-Figure1-1.png",
        "caption": "Figure 1. Kinematic model of crawler robot",
        "texts": [
            " KINEMATIC MODEL OF THE MASTER-SLAVE This paper studies crawler mobile robots, assuming: 1) the robot\u2019s left and right crawler wheels are exactly the same; 2) segment midpoint posed by two center of the circle and the straight line posed by robot\u2019s centroid connection are in the same direction with the robot speed; 3) the ground load evenly distributed; 4) two robot crawler is not slipping; 5) the exercise is two-dimensional movement in the horizontal plane. Consider the system composed of three robots, and the coordinates of the location of each robot are shown in Figure 1.The configuration of the main robot and the vice robots in the cartesian coordinate system is indicated by(xi,yi, i). Supposing the projection coordinates of robot's centroid Ci in the XOY reference plane is (xi,yi, i), where i=1,2,3, the attitude angle i of the robot stands for the angle between the running direction of robots and XOY reference plane X-axis positive. The robot position and orientation can be expressed by the vector P=[ xi,yi, i].The width of robot body is D, and the width of the crawler is b"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003560_s1070427211110085-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003560_s1070427211110085-Figure1-1.png",
        "caption": "Fig. 1. Calibration plots for the electrode based on a quinine\u2013 hydroquinone series system. (E) Potential, mV. Residence time in the buffer solution with pH 4.01, h: (1) 0, (2) 6, and (3) 71.",
        "texts": [
            " It was noted that the presence of fl uoride ions does not affect the potential of a quinhydrone electrode, while the infl uence is exerted by redox systems. According to data from [11], the theoretical function of quinhydrone electrodes is virtually unaffected by redox systems when in concentrations under 5 \u00d7 10\u20134 g-equiv l\u20131 if their oxidation potentials in examined solutions differ from that of the quinine\u2013hydroquinone system by no greater than 100 mV. The potential of a quinhydrone electrode varies with pH as 2.3RT aC6H4O2 2.3RT Eqh = E\u00b0qh + \u2013\u2013\u2013\u2013\u2013 log \u2013\u2013\u2013\u2013\u2013\u2013\u2013\u2013 \u2013 \u2013\u2013\u2013\u2013\u2013\u2013\u2013 pH. 2F aC6H4(OH)2 F Figure 1 shows the calibration plot (curve 1), averaged over the values obtained for several quinhydrone electrodes with standard buffer solutions on different days. The experimental value \u2202E/\u2202pH = 58 mV is close to theoretical; the standard electrode potential E\u00b0qh exp was well reproducible and was estimated at 490 \u00b1 2 mV vs. saturated silver chloride electrode [E\u00b0sc (20\u00b0C) = RUSSIAN JOURNAL OF APPLIED CHEMISTRY Vol. 84 No. 11 2011 201 mV], which corresponds to the data reported in [12] (E\u00b0qh = 696 \u00b1 1 mV)",
            " Good reproducibility of the E\u00b0 values allows interchangeability of electrodes fabricated by the same technology and their calibration using the same solution with a known pH value. Compared to an antimony electrode, a quinhydrone electrode exhibits a hydrogen function over a broader pH range; these electrodes can be used at pH < 8. In solutions with higher pH values the response of such electrodes is impaired because of the specifi c properties of the potential-forming system: Hydroquinone is a weak acid which dissociates via proton detachment (pK1 = 9.8, pK2 = 11.4 [12]). Curves 2 and 3 in Fig. 1 show that a 6-h residence of the electrode in solution leads to impaired response: (\u2202\u0415/\u2202pH)exp decreases to 54 mV, and E\u00b0qh exp, to 474 mV. After 71 hours of residence these values signifi cantly changed: \u2202E/\u2202pH decreased to 23 mV, and E\u00b0qh exp, to 383 mV. Thus, the response of the quinhydrone electrode gets negligible. The impaired response observed upon prolonged residence of electrode in solution is due to a faster elimination of the reduced (hydroquinone) species from the redox system compared to the oxidized (quinone) species [11]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002693_amr.605-607.175-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002693_amr.605-607.175-Figure2-1.png",
        "caption": "Fig. 2 Sleeve force analysis Fig. 3 Collet force analysis",
        "texts": [
            " While the crevice direction of the sleeve on the platform is random, the driven sprockets approach to the drive one (shown in Fig. 1a) until the sleeve is clamped (shown in Fig. 1b) when the drive sprocket begin to rotate driving the sleeve rolling under the friction action. The sleeve crevice is recognized by a photoelectric sensor, when it is found, the drive sprocket stops rolling while the driven one begin to leave and the sleeve also leave in the collet (shown in Fig. 1c), therefore the sleeve will fall in to the next feeding tube. Drive and driven sprocket design. Fig. 2 shows the force analysis of the sleeve in the process of champing. When the sleeve is champing by the driven sprockets, the drive sprocket rotates driving the sleeve rolling under the friction action, at this time the sleeve is applied f1M , 1F from the drive sprocket, 2F , 3F , f2M , f3M from the driven sprockets and 1M from the sliding friction between the sleeve and the platform. Formulation 1 is obtained on the basis of coplanar system of concurrent forces to be in equilibrium and the force of rolling friction equation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002470_amm.397-400.589-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002470_amm.397-400.589-Figure7-1.png",
        "caption": "Fig. 7 Equivalent stress contour of the outer ring",
        "texts": [
            " Comparison between FEA results and analytical solutions is shown in Table 1: Table 1 shows that analysis results of these two methods are close and simulation results are in good agreement with analytical solutions .The method that building a FEA model of rolling bearing and then analyzing its dynamic characteristics with ANSYS / LS-DYNA is feasible. Analysis of stress Fig. 4 is slice stress cloud of the bearing at 26ms. Fig. 5 is equivalent stress contour of rolling elements at 26ms. Fig. 6 is equivalent stress contour of the inner ring at 26ms. Fig. 7 is equivalent stress contour of the outer ring at 26ms. Fig. 4 Slice stress cloud of the bearing Fig. 5 Equivalent stress contour of rolling elements Fig. 4 shows equivalent stresses of the bearing are concentrated in the contact areas between the rolling elements and the inner/outer ring. The maximum stresses appear in a certain depth below the contact surfaces and gradually attenuate outwards. Hertz contact theory accounts that: there are compressive stresses in the elliptical contact areas of ball bearings; compressive stresses change in axial and normal directions on the surface; the maximum stresses appear in a certain depth of the contact regions[5]. In summary, simulation results in this paper are consistent with Hertz theory. It can be seen from Fig. 5 to Fig. 7, the maximum equivalent stresses of the inner ring, outer ring and rolling elements are not equal at the same time. Of all these three, rolling elements have a maximum stress of 312.9MPa, followed by the inner ring 106.9MPa and outer ring 84.7MPa. Relatively larger stresses of the rolling elements occur in the contact areas and the stresses in load zone are greater than that in non-load zone. Inner and outer rings directly contact with the rolling elements and bear most of the loads, whose stress situations influence the stress distribution of rolling elements greatly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001268_3-540-34319-9_8-Figure8.9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001268_3-540-34319-9_8-Figure8.9-1.png",
        "caption": "Figure 8.9: Omni-3",
        "texts": [
            " Inverse kinematics The inverse kinematics is a matrix formula that specifies the required individual wheel speeds for given desired linear and angular velocity (vx, vy, ) and can be derived by inverting the matrix of the forward kinematics [Viboonchaicheep, Shimada, Kosaka 2003]. 8.4 Omni-Directional Robot Design We have so far developed three different Mecanum-based omni-directional robots, the demonstrator models Omni-1 (Figure 8.8, left), Omni-2 (Figure 8.8, right), and the full size robot Omni-3 (Figure 8.9). The first design, Omni-1, has the motor/wheel assembly tightly attached to the robot\u2019s chassis. Its Mecanum wheel design has rims that only leave a few millimeters clearance for the rollers. As a consequence, the robot can drive very well on hard surfaces, but it loses its omni-directional capabilities on softer surfaces like carpet. Here, the wheels will sink in a bit and the robot will then drive on the wheel rims, losing its capability to drive sideways. \u00b7 FL \u00b7 FR \u00b7 BL \u00b7 BR 1 2 r -------- 1 1 d e 2 1 1 d e 2 1 1 d e 2 1 1 d e 2 vx vy Driving Program The deficiencies of Omni-1 led to the development of Omni-2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003815_amm.121-126.2211-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003815_amm.121-126.2211-Figure1-1.png",
        "caption": "Fig. 1 Three different crack opening modes. Fig.2 Model I of crack",
        "texts": [
            "33, Universidad Politecnica de Valencia, Valencia, Spain-26/05/15,11:35:16) Fracture Mechanics and the Basic Form of Crack. Fracture mechanics is a discipline that studies strength of crack and crack propagation law. It is a development and complement of the conventional design and has its own specific conditions. Usually the ways of crack expansion in the external force can be divided into three forms [13]: (1). model(I): opening in tension, (2). model(II): in-plane shear, (3). model(III): transverse shear. Three different crack opening modes are shown as Fig.1. If the crack is also affected by normal stress and shear stress simultaneously, that is to say: the model I and II exist at the same time, this is called composite crack. For the involute gear, the model of crack is composite. For two-dimensional equivalent model, the models of composite crack are I and II, and for three-dimensional model actually, there may be I, II and III models. For the spur gear, the model is usually equivalent to two-dimensional to calculate[14]. Fracture toughness and SIF"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001531_s00707-013-0995-y-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001531_s00707-013-0995-y-Figure5-1.png",
        "caption": "Fig. 5 Two contiguous elastic beams",
        "texts": [
            " This enables also to consider non-holonomic constraints \u00e0 priori (then, s\u0307 \u2208 IRg, g < f ). Remark The considered system is not restricted to a chain topology (as considered here). In case of branchings, all Fk within the rows of Eq. (19), which do not lie on the path, are equal to zero. In case of closed kinematic loops, we recommend to insert the corresponding constraint forces along with a Baumgarte [1] and Ostermeyer [19] stabilization. 3.2 Flexible beam systems The general case is depicted in Fig. 5. The procedures for purely elastic beams (Table 2) and for purely rigid bodies (Table 3) are now combined for the general case. The vector of describing velocities is known with Eq. (4) y\u0307 = ( vox voy \u03c9oz v\u0307 v\u0307\u2032 v\u0307\u2032\u2032 )T , (21) and the kinematic chain y\u0307i = Ti p y\u0307p + y\u0307i,rel reads \u239b \u239c\u239c\u239c\u239c\u239c\u239d vox voy \u03c9oz v\u0307(x, t) v\u0307\u2032(x, t) v\u0307\u2032\u2032(x, t) \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 i = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 Ai p \u23a1 \u23a31 0 \u2212vL 0 0 0 0 1 L 1 0 0 0 0 1 0 1 0 \u23a4 \u23a6 O \u2208 IR3,6 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 p \u239b \u239c\u239c\u239c\u239c\u239c\u239d vox voy \u03c9oz v\u0307L v\u0307\u2032 L v\u0307\u2032\u2032 L \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 p + \u239b \u239c\u239c\u239c\u239c\u239c\u239d vx,rel vy,rel \u03c9z,rel v\u0307(x, t) v\u0307\u2032(x, t) v\u0307\u2032\u2032(x, t) \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 i (22) where Ai p = \u239b \u239d cos \u03b3i p sin \u03b3i p 0 \u2212 sin \u03b3i p cos \u03b3i p 0 0 0 1 \u239e \u23a0 (23) (notice \u03b3i p = v\u2032 p L + \u03b3i see Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.33-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.33-1.png",
        "caption": "Fig. 2.33 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RPPaP (a) and 4RPPPa (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology R\\P||Pa\\kP (a) and R||P\\P||Pa (b)",
        "texts": [
            "29a) P\\R||P\\Pa (Fig. 2.21g) Idem No. 5 8. 4PRPaP (Fig. 2.29b) P\\R\\Pa \\kP (Fig. 2.21h) Idem No. 5 9. 4PPPaR (Fig. 2.30a) P\\P\\kPa\\kR (Fig. 2.21i) Idem No. 4 10. 4PPPaR (Fig. 2.30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002352_amm.404.194-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002352_amm.404.194-Figure3-1.png",
        "caption": "Fig. 3 The first four vibration motions of the ram slider and spindle tooling stage.",
        "texts": [
            " In addition, harmonic analysis was performed so as to examine the dynamic response of the spindle head when subjected to an external force, which was applied at the spindle tail simulating the cutting force in a frequency range. By this approach, we can examine the variation in the dynamic stiffness of the spindle tooling system and the influence of the preload of linear guides. Vibration modes. According to the finite element simulation, the first four vibration motions of the ram slider and spindle tool are illustrated in Fig. 3. In this case, the spindle head was equipped with the linear guides having low preload and was positioned at the lowest position. The four fundamental modes are associated with the first and second yawing vibrations of the spindle feeding stage, which are dominated by the linear guide modulus in Z axis. The other two modes are the first and second pitching vibrations of the spindle feed stage, with little part of the bending motion of the feeding stage of spindle head. It is obvious that the vibration frequencies corresponding to these modes are dominated by the linear guide modulus because they actually act as the support of the feeding stage"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001268_3-540-34319-9_8-Figure8.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001268_3-540-34319-9_8-Figure8.2-1.png",
        "caption": "Figure 8.2: Mecanum wheel designs with rollers at 90\u00b0",
        "texts": [
            " A Omni-Directional Robots 8 There are a number of different Mecanum wheel variations; Figure 8.1 shows two of our designs. Each wheel\u2019s surface is covered with a number of free rolling cylinders. It is important to stress that the wheel hub is driven by a motor, but the rollers on the wheel surface are not. These are held in place by ball-bearings and can freely rotate about their axis. While the wheels in Figure 8.1 have the rollers at +/\u2013 45\u00b0 and there is a left-hand and a right-hand version of this wheel type, there are also Mecanum wheels with rollers set at 90\u00b0 (Figure 8.2), and these do not require left-hand/right-hand versions. A Mecanum-based robot can be constructed with either three or four independently driven Mecanum wheels. Vehicle designs with three Mecanum wheels require wheels with rollers set at 90\u00b0 to the wheel axis, while the design we are following here is based on four Mecanum wheels and requires the rollers to be at an angle of 45\u00b0 to the wheel axis. For the construction of a robot with four Mecanum wheels, two left-handed wheels (rollers at +45\u00b0 to the wheel axis) and two right-handed wheels (rollers at \u201345\u00b0 to the wheel axis) are required (see Figure 8"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure7-1.png",
        "caption": "Figure 7. Comparison of Normalised stresses between without sub-modeling & with sub-modeling",
        "texts": [
            " Other than the identified critical regions are modeled with super elements using sub-modeling technique available in MSC / Nastran. These super elements are reduced to stiffness and load vector to interface DOF of these critical regions. Similar linear static analysis is carried out for identified critical regions along with reduced stiffness and load vector and compared the results with and without super elements. Boundary conditions for the identified critical regions are captured from super element without any user intervention. Figure 6 and Figure 7 shows the comparison of displacements and stresses respectively without and with sub-modeling. Displacement and stress results with sub-modeling is compared with full model linear static analysis and observed the same results. This has given confidence that submodeling is properly done. FE model should be modeled without any penetration. This is a pre-requisite to carry out contact analysis otherwise it will lead to unrealistic results. Figure 8 shows the super element at RSFB region and Figure 9 shows the Penetration plot at RSFB region"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003914_s1068798x13030052-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003914_s1068798x13030052-Figure3-1.png",
        "caption": "Fig. 3. Profiles of the tool\u2019s rear surface with different front surface geometry, when machining steel 45 by means of a T15K6 alloy tool: v = 3.3 m/s; s = 0.4 mm/turn; cut ting depth 2 mm.",
        "texts": [
            " Plastic deformation of the tool\u2019s cutting edge may be accompanied by the appearance of flaws due to creep. Hence, the creep rate will affect the tool\u2019s time of flaw free operation (its life). It is important here to establish the factors that affect the creep rate and to find means by which it may be reduced. The geometry of the tool\u2019s cutting section has con siderable influence on the creep rate, which declines with decrease in the rear angle at the primary cutting edge. The creep rate is lowest for tools with the front surface geometry in Fig. 3. The facet width ff = 0.18 mm [2]. The creep rate declines with increase in radius at the cutter tip (especially in the section adjacent to cut ter tip) and with decrease in the primary plane angle (Fig. 3). (1) The creep rate increases with decrease in ther mal conductivity of the hard alloy. That is associated with displacement of the isotherms corresponding to creep toward the tool\u2019s cutting edge\u2014that is, toward smaller tool cross sections. (2) The creep rate increases with the cobalt content in the hard alloy, as a result both of the greater content 325\u00b0C 655 640 525 460 410\u00b0C 575\u00b0C 615\u00b0C 530\u00b0C 300\u00b0C 610585565 485430 300\u00b0C (a) (b) Fig. 2. Temperature field of the tool\u2019s cutting edge in machining steel 45 by means of a VK8 alloy (a) and T15K6 alloy (b) tool: v = 1 m/s; s = 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001869_9781118530009.ch9-Figure9.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001869_9781118530009.ch9-Figure9.1-1.png",
        "caption": "FIGURE 9.1 (a) Schematic of CE system coupled to off-column AD: (a) buffer reservoir; (b) separation capillary; (c) detection capillary; (d) eluent. (b) Detailed schematic of porous joint: (a) microscope slide; (b) fused silica capillary; (c) porous glass capillary; (d) joint; (e) epoxy; (f) polymer coating.",
        "texts": [
            " The authors carefully constructed a fracture at the cathodic end of the separation capillary and then joined both parts of the capillary with porous glass capillary to create an electrically conductive porous glass joint that decouples the electrochemical detector from the electrophoretic separation current. The electrophoretic separation potential dropped across the capillary prior to the porous joint and the resulting electroosmotic flow acted as a \u201cpump,\u201d pushing both buffer and solute bands through the short section of capillary after the joint to the EC detector, as illustrated in Figure 9.1. This section of the capillary had ground potential applied so that there was very little interference from the high potential field on the electrodes used in the detection cell. This AD detector consisted of a carbon fiber WE of 10mm diameter and 0.1\u20131mm length, with a saturated calomel RE and stainless steel CE. Careful construction of the glass joint and improvements in design reduced band broadening between the decoupler and the detector to the point where it was considered negligible [10\u201312]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001230_0470036427.ch3-Figure3.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001230_0470036427.ch3-Figure3.16-1.png",
        "caption": "Figure 3.16 Phasor addition.",
        "texts": [
            " Thus, in order to multiply the I and Z phasors together, we draw a new vector at an angle with the x axis corresponding to the sum of the two angles, and with a length corresponding to the product of lengths. Note that because this product has different physical dimensions or units (voltage) than its factors (current and impedance), the new arrow can be drawn on the same diagram, but its length cannot be compared to phasors representing other units. If two or more phasors are to be added together\u2014which is done graphically with amazing ease by placing them end to end\u2014they must represent quantities of the same dimension drawn to the same scale.21 Consider, for example, the a.c. circuit in Figure 3.16. The phasor diagram is drawn in terms of voltage. It simultaneously shows several things about the circuit: First, it illustrates that the total voltage drop around the circuit adds up to zero, as required by Kirchhoff\u2019s voltage law (KVL) (see Section 2.3.1). It does this very simply by placing voltage phasors head to tail, adding them just like voltage drop is added in series along the circuit path, and requiring that the destination or sum correspond to the voltage drop measured around the other side of the circuit",
            " If we allow all the phasors to rotate and plot their projection onto the horizontal (real) axis, they generate sine waves for the respective voltages. These sinusoidally varying voltages IR and IX with their distinct magnitude and phase could be physically measured across each circuit element if each element modeled here in fact corresponded to a distinct physical object, rather than one abstracted property (resistance or reactance) of a real object (like a wire coil) that actually manifests both properties in the same space. The phasor diagram in Figure 3.16 offers a visual explanation of how the phase angle or power factor of a circuit is determined by the combination of its elements. Note that each individual element could be drawn in any arbitrary direction\u2014for example, were we dealing with a resistance alone, we probably would have chosen to make it horizontal\u2014but that the combination of elements in the circuit dictates the phasors\u2019 relationship to each other. Thus, the relative magnitude of resistance and reactance in the same circuit determines the proportions of the triangle and the angle between voltage and current"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002441_amr.544.286-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002441_amr.544.286-Figure8-1.png",
        "caption": "Fig. 8 Mesh models for the present FE analysis",
        "texts": [
            " Dissembled the linking rods and the side frames, one wheelset, one vice-frame, and two pieces of rails forms a merged part with clear dynamic balance force system as shown as in Fig. 7, containing additionally wheel-rail contacts and interference fits between two cambered surfaces of the vice-frame and outer surfaces of two bearings. Finite Element (FE) Analysis. An elastic-plastic FE analysis was performed to obtain service stresses of the bearing based on a piece of inspected service load spectra of the car [5]. Integrated FE modeling was applied to the merged part and a policy of two stages was employed for the calculation process. As shown as in Fig. 8, in stage I vice-frame, wheels, axle, and two pieces of rails are physically analyzed together on a basis of taking bearing as one (outer ring-rollers-inner ring) merged solid. And then, in stage II the local part around the bearing including additionally pieces of axle and vice-frame is calculated in the manner of taking the stress-stain parameters on the cut sections as constraint conditions and the bearing three parts, i.e. outer ring, rollers and inner ring, as contacting physically. Results and Discussions"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002047_978-1-4471-5110-4_6-Figure6.15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002047_978-1-4471-5110-4_6-Figure6.15-1.png",
        "caption": "Fig. 6.15 Problem 6.2",
        "texts": [],
        "surrounding_texts": [
            "If curl P = \u2207\u00d7P = 0 the force P is conservative. The expression \u2207\u00d7P is calculated with MATLAB: rotP_ = curl(P_,[x y z]); and the results is: curl(P_)=[0, 0, 0] The potential energy is calculated with V (x, y, z) = \u2212 \u222b P \u00b7 dr + C = \u2212 \u222b [\u2212kxdx \u2212 kydy + (k R \u2212 kz \u2212 mg)dz] + C V (x, y, z) = k x2 2 + k y2 2 + k (R \u2212 z)2 2 + mgz + C, (6.8) where C is an arbitrary constant known as the constant of integration. The equation of the sphere can be written as x2 + y2 + z2 = R2 or z = \u221a R2 \u2212 x2 \u2212 y2, (6.9) and with Eq. (6.8) the potential energy function of x and y is calculated in MATLAB with: V1=-int(P_(1)); V2=-int(P_(2)); V3=-int(P_(3)); V=V1+V2+V3+C; fVxy=subs(V,z,sqrt(R\u02c62-x\u02c62-y\u02c62)); Vxy=simple(simplify(Vxy)); and V (x, y) is obtained as V (x, y) = k R2 + (mg \u2212 k R) \u221a R2 \u2212 x2 \u2212 y2 + C. The partial derivative of the function V (x, y) with respect to x and y are \u2202V \u2202x = \u2212 (mg \u2212 k R) x\u221a R2 \u2212 x2 \u2212 y2 , \u2202V \u2202y = \u2212 (mg \u2212 k R) y\u221a R2 \u2212 x2 \u2212 y2 . The equilibrium positions of the particle are obtained from \u2202V \u2202x = 0, \u2202V \u2202y = 0. In MATLAB the equilibrium positions are obtained with: dVxydx = simple(diff(Vxy,x)); dVxydy = simple(diff(Vxy,y)); xe=solve(dVxydx,x); ye=solve(dVxydy,y); ze=solve(xe\u02c62+ye\u02c62+z\u02c62-R\u02c62,z); The results for the equilibrium positions are M1(0, 0, R) and M2(0, 0,\u2212R). Example 6.5 A particle P of mass m is on a circle of radius R as shown in the Fig. 6.13. The circle is on a vertical plane xy. Find the equilibrium positions of the particle. Solution The independent variable is the angle \u03b8. The position of the particle P is x = R*cos(theta); y = yN+R+R*sin(theta); r_ = [x y]; where yN is the y coordinate of the lower end N of the circle. The gravity is the only force acting on the particle and the potential energy is calculated with: dr_=diff(r_,theta); G_ = [0 -m*g]; V = -int(G_*dr_.\u2019); fprintf(\u2019V=%s + C\\n\u2019, char(V)) The MATLAB expression for the potential energy is: V=R*g*m*sin(theta) + C where C is a constant of integration. The equilibrium positions are calculated from the equation: dV = diff(V,theta); thetae=solve(dV,theta); theta1=thetae; theta2=theta1+pi; The equilibrium position are the points M and N as shown in Fig. 6.13: theta1 = pi/2 and theta2 = (3 \u2217 pi)/2. The equilibrium stability is verified with the second derivative of the potential energy: d2V = diff(dV,theta); d2V1=subs(d2V,theta,theta1); d2V2=subs(d2V,theta,theta2); and the MATLAB results are: d2V/d(theta)\u02c62=-R*g*m*sin(theta) for theta1 => d2V/d(theta)\u02c62=-R*g*m for theta2 => d2V/d(theta)\u02c62=R*g*m The equilibrium position \u03b8 = 3 \u03c0/2, position N , is a stable equilibrium because d2V/d\u03b82 = Rgm is positive. 6.6 Problems B A C 1 2 k \u03b8 m, l m, l F A B A C \u03b8 m, l l k k Fig. 6.18 Problem 6.5 B A k O C F \u03b8 Fig. 6.19 Problem 6.6 B A k \u03b8 m, l k Fig. 6.21 Problem 6.8 \u03b8 l m, l m, l m, l F B A C D k 1 2 3 0 0 by the torsion spring is M = K \u03b8, where \u03b8 is the relative angle between the links at the joint. Determine the minimum value of K which will ensure the stability of the mechanism for \u03b8 = 0. 6.8 Figure 6.21 shows a four-bar mechanism with AD = l. Each of the links has the mass m (m1 = m2 = m3 = m) and the lenght l (l1 = l2 = l3 = l). At B a vertical force F acts on the mechanism and the spring stiffness is k. The motion is in the vertical plane. Find the equilibrium angle \u03b8. Use the following numerical application: l = 15 in, m = 10 lb, F = 90 lb, and k = 15 lb/in. Select an unextended (initial) length L0 for the spring. 6.9 A particle of mass m can move freely in space. The potential energy V of the particle at x = l, when the particle is subject to a vertical force F = ax2 + bx + c, is V = s. Find the equilibrium positions of the particle. For the numerical application use a = 1, b = \u22123, c = 0, l = 0 m, and s = 5 J. 6.10 A bar of mass m and length l is supported by a vertical wall and a point at O , as shown in Fig. 6.22. Find the equilibrium positions of the bar. For the numerical application use l = 0.5 m, a = 0.1 m, m = 1 kg, and g = 9.81 m/s2."
        ]
    },
    {
        "image_filename": "designv11_100_0003653_iceee.2012.6421204-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003653_iceee.2012.6421204-Figure3-1.png",
        "caption": "Figure 3. The forces acting on the UAV",
        "texts": [],
        "surrounding_texts": [
            "The mathematical model for a rigid body [2], [14] expressed in the body frame for the translational and rotational dynamics is: m _V B +mV B !B = FB +mGB (10) I _!B + I!B !B = B (11) 1) Translational dynamics: The translational dynamics fol- lows rewriting (10) in the inertial frame: m _V I = R1I!B 1 FB +mGI (12) (12) can be rewritten as: m x = FBx (c c ) F B y (c s ) + F B z (s ) (13) m y = FBx (c s + s s c ) + F B y (c c s s s ) FBz (s c ) (14) m z = FBx (s s c s c ) + F B y (s c + c s s ) +FBz (c c ) mg (15) The generalized forces are described in gure 3 and are given next. The general vector of forces is: FB = FBx FBy FBz T = FBA + F B M T : (16) where the generalized vector of motors forces in the body frame is de ned as: FBM = RM!B 5 FMM1 +RM!B 5 FMM2 + FBM3 + FBM4 T (17) and the generalized vector of aerodynamic forces in the body frame is de ned as: FBA = [RS!B 4 FSAwb +R S!B 4 FSAt +R S!B 6 FSAr] (18) Then the translational dynamics of the UAV in explicit form is: m x = (c c ) ( cBcCDwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) sB(Lwb1(B; Ea1) + Lwb2(B; Ea2) + Lt(B; Ee)) cCDr(C; Er) sCLr(C; Er) + FM1s + FM2s ) (c s ) ( sC (Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) +sCDr(C; Er) cCLr(C; Er)) + (s ) ( cCsB (Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) +cB(Lwb1(B; Ea1) + Lwb2(B; Ea2) + Lt(B; Ee)) +FM1c + FM2c + FM3 + FM4) (19) m y = (c s + s s c ) ( cBcC(Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) sB(Lwb1(B; Ea1) + Lwb2(B; Ea2) +Lt(B; Ee)) cCDr(C; Er) sCLr(C; Er) +FM1s + FM2s ) + (c c s s s ) ( sC(Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) + sCDr(C; Er) cCLr(C; Er)) (s c ) ( cCsB (Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) +cB(Lwb1(B; Ea1) + Lwb2(B; Ea2) + Lt(B; Ee)) +FM1c + FM2c + FM3 + FM4) (20) m z = (s s c s c ) ( cBcC(Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) sB(Lwb1(B; Ea1) + Lwb2(B; Ea2) +Lt(B; Ee)) cCDr(C; Er) sCLr(C; Er) +FM1s + FM2s ) + (s c + c s s ) ( sC(Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) + sCDr(C; Er) cCLr(C; Er)) + (c c ) ( cCsB (Dwb1(B; Ea1) +Dwb2(B; Ea2) +Dt(B; Ee)) +cB(Lwb1(B; Ea1) + Lwb2(B; Ea2) + Lt(B; Ee)) +FM1c + FM2c + FM3 + FM4) mg (21) 2) Rotational dynamics: From (11) the rotational dynamics of the UAV in the body frame is: I _!B + I!B !B = B (22) Expressing the angular speed !B on the Euler's angles [1], [2], [3]: !B = R _q2 (23) Using the Euler angles (roll, pitch, yaw) the rotational matrix is: R = 0 @ c c s 0 s c c 0 s 0 1 1 A (24) We take the rst derivate of (23) to obtain: _!B = _R _q2 +R q2 (25) Solving (22) for _! we get: _!B = I 1 B I!B !B (26) Developing both sides of the equation and solving for q2 = ; ; : = 1 c c ( s + _ _ c s + _ _ s c _ _ c + 1 Ix Bx qr (Iy Iz) ) (27) = 1 c ( s c _ _ s s + _ _ c c +_ _ s + 1 Iy By pr (Iz Ix) ) (28) = s _ _ c + 1 Iz Bz pq (Ix Iy) (29) The generalized vector of couples is de ned as: B = Bx By Bz T = BM + BA + B D + B G T : (30) In a similar way we obtain the rotational dynamic of the UAV: = 1 c c s + _ _ c s + _ _ s c _ _ c + 1 Ixc c (c (yM1FM1 y2MF2M ) + y3MF3M y4MF4M + sCzwb(Dwb1(B; Ea1) +Dwb2(B; Ea2)) +cCzrLr(C; Er) sCzrDr(C; Er) + s ( DM1 DM2) + c qIzM (!M2 !M1) qr (Iy Iz))(31) = 1 c s c _ _ s s + _ _ c c + _ _ s + 1 Iyc ( c (xM1FM1 + x2MF2M ) + x3MF3M +x4MF4M sBzwb (Lwb1(B; Ea1) + Lwb2(B; Ea2)) cBcCzwb(Dwb1(B; Ea1) +Dwb2(B; Ea2)) +cBxtLt(B; Ee) cCsBxtDt(B; Ee) sCzrLr(C; Er) cCzrDr(C; Er) + c pIzM (!M1 !M2) +s rIzM (!M1 !M2) pr (Iz Ix)) (32) = s _ _ c + 1 Iz (s (yM1FM1 y2MF2M ) +sCxtDt(B; Ee) + cCxrLr(C; Er) sCxrDr(C; Er) + c ( DM2 DM1) + DM3 DM4 + s qIzM (!M2 !M1) pq (Ix Iy)) (33) III. CONTROL"
        ]
    },
    {
        "image_filename": "designv11_100_0003882_icdma.2012.156-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003882_icdma.2012.156-Figure8-1.png",
        "caption": "Figure 8. Model of the gear",
        "texts": [
            " In the above steps, each size is parameterized and written in VB language. Figure 5 is the main interface of the software. IV. FAST MODEL-BUILDING EXAMPLE Known a national standard double circular gear module (Mn=8), helix angle ( \u03b2=15 ), gear teeth number ( z=50 ), teeth width ( b=120 ). (1) Input these known parameters to the interface window as shown in Figure 6. (2) Click \"Generate\" button to display the reference radius, root radius, tip radius shown in Figure 7 and generate three-dimensional solid model shown in Figure 8. V. CONCLUSION In this paper the solid model of the double circular-arc gear is quickly generated on the basis of the double circulararc gear end-face teeth profile equation and the use of CATIA secondary development platform. Main achievements are as follows: (1) Having realized parameterized modeling of the gears, users only need to enter gear teeth number ( z ), modulus ( Mn ), helix angle ( \u03b2 ), teeth width ( b ), then a 3D solid model of the whole gear can be quickly generated, which is easy to use and solves the problem of low efficiency of the double arc gear design"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.41-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.41-1.png",
        "caption": "Figure 12.41 A link (i) and its force system.",
        "texts": [
            "634) \u03d5\u0307 = \u23a7 \u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 \u03c93I3 2I1 cos \u03b8 ( 1 + \u221a 1 \u2212 4glmI1 \u03c92 3I 2 3 cos \u03b8 ) cos \u03b8 = 0 gl m \u03c93I3 cos \u03b8 = 0 (12.635) 12.5 MULTIBODY DYNAMICS We consider multibodies as a set of rigid bodies connected to each other by revolute or prismatic joints. Most of the mechanical devices are multibodies. Each body is called a link. We follow the method of assigning numbers and coordinate frames of links as described in Chapter 7. Figure 12.40 illustrates a link (i) of a multibody along with the velocity and acceleration vectorial characteristics. Figure 12.41 illustrates a free-body diagram of link (i). The force 0Fi\u22121 and moment 0Mi\u22121 are the global expressions of the resultant force and moment that link (i \u2212 1) applies to link (i) at joint i. Similarly, 0Fi and 0Mi are the global expressions of the resultant force and moment that link (i) applies to link (i + 1) at joint i + 1. We measure and show the force systems (0Fi\u22121, 0Mi\u22121) and (0Fi , 0Mi) at the origin of the coordinate frames Bi\u22121 and Bi , respectively. The sums of the external loads acting on link (i) are shown by \u2211 0Fei and \u2211 0Mei "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003848_imece2012-87624-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003848_imece2012-87624-Figure5-1.png",
        "caption": "FIGURE 5. TRANSFORMATION OF JOINT k",
        "texts": [
            " The internal forces he i due to deformation are given by he i = Kiqi +Diq\u0307i = [ 0 0 0 Ke i ] [ xI i qi ] + [ 0 0 0 De i ] [ xII i q\u0307i ] , (2) where Ke i is the elastic stiffness matrix and De i is the elastic damping matrix, which are part of the standard data set [4]. The Newton-Euler equations of every body Ki i = 0 . . . n within an FMBS system with n bodies can be written as[ MxII i Mtr iT Mtr i Me i ] \ufe38 \ufe37\ufe37 \ufe38 :=Mi [ \u02d9xII i q\u0308i ] = he i + hQ i. (3) The joint k connects the node of M i,src of body Ki with the node Mk,dst of body Kk as depicted in Figure 5 and [6, pg. 245] as well as [16]. The position of the reference frame of bodyKi is given by pi, the orientation by the rotation matrixAi, in addition the translational velocities are p\u0307i and the rotational velocities are \u03c9i. Therefore the position, orientation and the translational and rotational velocity of the node M i,src can be evaluated pi,src = \u0398i,src ( pi + R\u0302 i,src) , (4a) Ai,src = \u0398i,srcAi, (4b) where \u0398i,src = \u0398(\u03a8i(Ri,src)qi) describes the rotation of a frame attached to node M i,src due to deformations of the body, Ri,src denotes the position of the node M i,src within the reference configuration of the body Ki and R\u0302 i,src = Ri,src + \u03a6i,srcqi denotes the position of the node within the current deformation of the body"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002042_9783527643981.bphot009-Figure2.4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002042_9783527643981.bphot009-Figure2.4-1.png",
        "caption": "Figure 2.4 Schematic representation of (a) hybrid sensor and (b) implanted hybrid sensor [26]. Reprinted fromA.Pasic,H. Koehler, I. Klimant, and L. Schaupp, Miniaturized fiberoptic hybrid sensor for continuous glucose monitoring in subcutaneous tissue, Sens.",
        "texts": [
            ", 1996, 68, 1408\u20131413. Copyright 1996 American Chemical Society. solution, or by compensating the variable oxygen background by determining it independently, for example, with a second sensor. This kind of dual sensor was reported first by Li and Walt [18], subsequently by Wolfbeis et al. [22], and more recently by Klimant s group [25, 26]. They used a dual sensor, consisting of two commercially available oxygen optodes in close proximity, where one had been modified with GOx and the other served as a reference (Figure 2.4). This sensor is unaffected by fluctuations of the oxygen concentration in the sample and also allows for the compensation of slight temperature fluctuations. Comparable sensor schemes have been patented [27\u201329]. 2.4.2 Sensors Based on the Use of Microparticles (mPs) and Nanoparticles (NPs) Particle-based sensors have attracted substantial interest for intracellular glucose testing in recent years. They may be applied in the future in the bloodstream as molecular analytical machines reporting blood glucose levels, provided that the optical signal they are giving canbe interrogated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003431_2013-36-0272-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003431_2013-36-0272-Figure2-1.png",
        "caption": "Figure 2 - Tire geometric elements",
        "texts": [
            " It means that the equation (1) is not valid, and we must find another way to express that relationship. One possible approach would be through the use of mathematical formulation with curvilinear coordinates and different metrics. Instead, we will make a physical approach, using geometric concepts to analyze the problem. Let us set the following general hypothesis: Page 2 of 7 1 - the sliding is negligible, i.e., the relative displacement between a point of the tire in contact with the ground, and the ground, is negligible. 2 - the tire is constituted by three elements, as shown in Figure 2: a cylindrical segment membrane of radius R, width L and thickness h \u2013 the tire belt \u2013 and two annular membranes \u2013 the sidewalls. 3 - the tire supports an internal pressure p, and all those membranes are initially in traction. 4 - these membranes are perfectly flexible, i.e., they have no bending strength. Their surfaces present a plane stress state. 5 - these membranes are elastic. i.e., they present inplane deformations. These deformations, however, are small and do not change the cylindrical shape of the tire belt region not in contact with the ground"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002567_icmtma.2011.343-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002567_icmtma.2011.343-Figure1-1.png",
        "caption": "Figure 1. The wire race ball bearing used in a rotating platform",
        "texts": [
            " Therefore, it has been extensively used in many fields characterized by both of little space and lightweight during recent years, such as large radars, aircraft simulating rotary tables and turrets, etc[1]. As a specific application, a wire race ball bearing applied in a rotating platform is designed depending on the needs of saving space and mass. Moreover, the bearing possesses the capability of sustaining loads and moments from whatever direction. The configuration picture of the wire race ball bearing used in the rotating platform is shown in Figure 1. However, the cyclic alternating stress between wires and balls, which is produced by load and impact in the course of practical work, often makes the contact bodies destroyed, especially for the wires[1]. This would affect the reliability of the whole system. Therefore, it is very important for the wire race ball bearing to make sure that the mechanical performance of rolling elements are realiable during operating. In this paper, to assure the reliability of the bearing, the contact stress and deformation between the ball and inner wire race are calculated repectively by the Hertz contact theory and finite element simulation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003448_amr.463-464.1304-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003448_amr.463-464.1304-Figure1-1.png",
        "caption": "Fig. 1. Kinematics scheme of two degree of mobility mechanisms",
        "texts": [],
        "surrounding_texts": [
            "first part are presented some kinematics schemes used to plane manipulators. These mechanisms are used to manipulators, positioning and control systems. In generally these mechanisms have two or three degree of mobility. The purpose of the paper is to study the dynamics of a plane manipulator mechanical system used to manipulate garbage containers. It is presented the kinematics scheme of a plane manipulator, used to this purpose and is presented the mechanism functional description. In the second part is presented the kinematical and dynamical analysis for the plane manipulator mechanism. In the last part of the paper are presented graphical results for the dynamics parameters.\nIntroduction\nThe aim of this research paper is to study the dynamic answer of a parallel manipulator, which is used to containers lifting. Also, is studied the dynamics of the mechanism, in the purpose of the actuation moment minimization.\nIn the literature, they are some studies concerning the dynamics and optimization of different\nmanipulators.\nIn generally, the mechanisms with hydraulic actuation are manipulation, positioning and control mechanisms. These mechanisms are in the structure of the harvester machine, industrial robots, and manipulators. The mechanisms mobility correspond with the number of independents parameters, necessary to assure the mechanism functionality. These mechanisms can be characterized by total degree of freedom, partial or fractioned [2, 3]. The actuation of those modules is realized with mechanical \u2013 hydraulics systems, type cylinder piston (CH).\nThe manipulator structure\nThe mechanism works sequentially, with a single motor element (the crank A0A). The kinematics\u2019 scheme of the plane mechanism [2, 3, 4], with two degree of mobility, is presented in figure 2.\nAll 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-20/05/15,22:12:43)",
            "The characteristics dimensions are, (in conformity with figure 2): ;380;680;100 660 mmlCDmmlCGmmeGA =\u2032=====\n.120;440;540;620\n;240;680;220\n5543\n3210\nmmlFHmmlEFmmlDEmmlCF\nmmlBCmmlABmmlAA\n=\u2032======\u2032=\n======\nThe maximum rotation angle of the crank shaft 1 is 260\u00b0. In the mechanism functioning we\nidentify two phases [4]:\n1. The element 6 (with the points G, C and D) stays fixed until, trough the crank shaft rotation, the\npoint B reach on the vertical part of the element 6, between the points C and G;\n2. All the kinematics elements of the mechanism are joining rigid, also continuing the crank shaft rotation until the end position, the mechanism become as a rigid body, which rotate upon the fixed joint A0. The trajectories of all mobile joints are circles with the centre in A0 joint.\nThe kinematics analysis of the manipulator mechanism\nFor the first kinematics chain ABC, we write the positions equations:\n   \u22c5+=\u22c5+ \u22c5+=\u22c5+ 322 322 cossin coscos \u03d5\u03d5 \u03d5\u03d5 CBCA CBCA lxly lxlx (1)\nAfter square up of equations (1), and summing, we obtain:\n3231\n22 2 2 1 2 2 sin2cos2 \u03d5\u03d5 \u22c5\u22c5+\u22c5\u22c5+++= elelleel CBCBCB where:\nAC xxe \u2212=1 ; AC yye \u2212=2 .\nWe obtain a trigonometrically equations with variable coefficients, under the form:\n0cossin 13131 =++ CBA \u03d5\u03d5 , where: 21 2 elA CB \u22c5\u2212= ; 11 2 elB CB \u22c5\u2212= ; 22 2 2 1 2 21 CBleelC \u2212\u2212\u2212=\nThe angle 3\u03d5 and 2\u03d5 are obtained by solution of the equations:\n  \n\n  \n\n\u2212\n\u2212+\u00b1 =\n11\n2 1 2 1 2 11\n3 2 CB\nCBAA arctg\u03d5 ;\n  \n\n  \n\n\u2212\n\u2212+\u00b1 =\n22\n2 2 2 2 2 22\n2 2 CB\nCBAA arctg\u03d5 (2)\nwhere: AC xxe \u2212=1 ; AC yye \u2212=2 ; 222 2 elA \u22c5\u2212= ; 122 2 elB \u22c5\u2212= ; 2 2 2 2 2 1 2 2 leelC CB \u2212\u2212\u2212= For the second kinematics chain DEF we write the equations for positions:\n  \n\u22c5+=\n\u22c5+=\n3\n3\nsin\ncos\n\u03d5\n\u03d5\nCFCF\nCFCF\nlyy lxx ,   \n\u22c5+=\n\u22c5+=\n5\n5\ncos\nsin\n\u03d5\n\u03d5\nEFFE\nEFFE\nlyy\nlxx , or   \n\u22c5+=\n\u22c5+=\n4\n4\nsin\ncos\n\u03d5\n\u03d5\nDEDE\nDEDE\nlyy\nlxx (3)\nAfter square up of equations (3), and summing, we obtain:\n4443\n22 4 2 3 2 sin2cos2 \u03d5\u03d5 \u22c5\u22c5+\u22c5\u22c5+++= elelleel DEDEDEEF\nThat is, an equation under the form: 0cossin 44444 =++ CBA \u03d5\u03d5 where: 44 2 elA DE \u22c5\u2212= ; 34 2 elB DE \u22c5\u2212= ; 22 4 2 3 2 4 DEEF leelC \u2212\u2212\u2212= ; FD xxe \u2212=3 ; FD yye \u2212=4"
        ]
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.61-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.61-1.png",
        "caption": "Figure 12.61 A turning uniform beam with a tip mass.",
        "texts": [
            " Variable Mass Moment Derive the general Euler equations of motion of a rigid body B with variable mass moment matrix [I ] = [I (t)]. 27. A Turning Uniform Link Figure 12.60(a) depicts a triangular link attached to the ground by a revolute joint at O. The free-body diagram of the link shows the gravity and the driving force and moment at the joint, as shown in Figure 12.60(b). Determine the Newton\u2013Euler equations of the triangular link. F0 Y y1 x1 X m1 Q0 Q0 m1g n m B1 B0 C \u03b8 l1 (a) (b) g Figure 12.60 A turning uniform triangular link. 28. A Turning Uniform Beam with a Tip Mass Consider the uniform massless beam of Figure 12.61(a) with a hanging mass m2 at the tip point. Figure 12.61(b) illustrates the free-body diagram of the beam. Determine the equations of motion of the beam. 29. 2R Planar Manipulator Newton\u2013Euler Dynamics A 2R planar manipulator and its free-body diagram are shown in Figure 12.62. The torques of actuators are parallel to the Z-axis and are indicated by Q0 and Q1. Use the multibody Newton\u2013Euler dynamics and determine the equations of motion to find the required torques Q0 and Q1 and the joint forces F0 and F1. F0 x2 y2 y0 y1 x1 x0 m1 m2 Q0 Q1 Q0 \u2212F1 F1 m1g m2g\u2212Q1 n1 m1 n 2 m 2 B1B0 C1 C2 \u03b82 \u03b81 l1 l 2 B2 Figure 12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure17-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure17-1.png",
        "caption": "Figure 17. Normalised Stress at RSMB region (Contact analysis)",
        "texts": [],
        "surrounding_texts": [
            "Surface to surface contact is defined for the following contact pairs \u2022 RSFB bracket & External flitch \u2022 External Flitch & FSM \u2022 FSM & Internal flitch \u2022 Internal Flitch & RSFB Cross member"
        ]
    },
    {
        "image_filename": "designv11_100_0002780_s10851-011-0286-y-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002780_s10851-011-0286-y-Figure6-1.png",
        "caption": "Fig. 6 Geometry of the leg\u2019s articulations in the Aibo robot. Lateral and frontal views",
        "texts": [
            " In order to determine which leg tips are the supporting points, we proceed as follows: (1) We compute the tip position of each leg relative to the body center coordinate reference system from the knowledge of the joint angles of each leg articulation. (2) We compute the hypothetical supporting planes defined by each combination of three leg tip points. (3) We discard hypothetical supporting planes for which at least one leg tip is below it or do not comply with the stability condition about the center of mass. First we compute the position of the leg tip g applying the transformations defined by the chain of articulations from the leg tip up to the body center, as shown in general in Fig. 5. Figure 6 shows the leg parameter specification for the Aibo robot. The In homogeneous coordinates, the leg tip is computed by the following product of elemental transfor- mation matrices: ( g 1 ) = (Tn+1 \u00b7 Rn.Tn \u00b7 \u00b7 \u00b7R1.T1). ( 0 1 ) , (13) where Rk is the rotation matrix corresponding to the k-th leg articulation from tip to the body center, being n the number of articulations and Tk\u22121 the translation matrix corresponding to the leg segment between the (k \u2212 1)-th and the k-th articulations. Translation matrix T1 corresponds to the translation from the tip to the first articulation, while translation matrix Tn+1 corresponds to the translation from the last articulation to the body center reference system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002563_ijmms.2013.052783-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002563_ijmms.2013.052783-Figure4-1.png",
        "caption": "Figure 4 A size-adjustable IET",
        "texts": [
            " The size of equilateral triangles to be formed is changed depending on the initial distribution of the robots or the trajectory of their locomotion. So, we extend the IET algorithm that is explained above to specify the size of an equilateral triangle to be formed. For example, let us consider the case of forming an equilateral triangle, whose one side is given by l. Although the distance d between other two robots was used to imagine an IET as described above, the segment of l length is here given to form an equilateral triangle with the desired size (Figure 4). Figure 4 shows the case of d > l, in which each robot may gather so as to make a smaller triangle, whereas if the case of d < l, then they may spread for making a larger triangle. Thus, letting one side length of IET be the same as that of a desired equilateral triangle, the group can make a desired formation. For example, consider a hexagonal formation as shown in Figure 1. Assume that there are seven robots and they are given the reference robots shown in Table 1. Assuming that an equilateral triangle formation is formed with the IET algorithm, three robots (Robot 1, Robot 2 and Robot 4) can form the equilateral triangle formation shown in Figure 5(a)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002737_09507116.2011.590664-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002737_09507116.2011.590664-Figure2-1.png",
        "caption": "Figure 2. Groove and torch settings.",
        "texts": [
            " However, if the carbon dioxide content is lower than 10%, there is an effect from the plasma flow and the penetration becomes rather finger shaped7,8. In such cases, there is a risk that the melting of the groove face either in the groove or, in T-joint welding, on both sides of the penetration bottom may be incomplete, resulting in poor fusion. To confirm that appropriate weaving conditions can be set and adequate melting of the groove face is possible, a Tjoint of 6 mm thick SM490A steel sheet base metal was lap welded with an intended leg length of 6.5 mm, with the groove and torch settings as shown in Figure 2, and the penetration dimensions were measured as shown in Figure 3. Figure 4 shows a photograph of the appearance of the weldment and a macrograph of its cross section. It is clear from Figure 4 that fusion of the groove root area is unproblematic when welding is carried out with Ar \u00fe 3.5% O2 two-component mixed gas, with a penetration depth similar to that achieved when Ar \u00fe 3.5% O2 \u00fe 20% CO2 and with a good bead appearance. Figure 5 shows the relationship between spatter loss, shielding gas, and arc voltage when the carbon dioxide is 20%; if the arc voltage is set low to prevent undercut, there is an increase in the spatter quantity whereas with Ar \u00fe 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002513_amm.164.497-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002513_amm.164.497-Figure1-1.png",
        "caption": "Fig. 1 Structure diagram",
        "texts": [
            " With the application of these two Cleaning machines, production efficiency is increased significantly. however, limited by the principle of the device structure, the spiral roller clearance is less efficient in dealing with the coal slime which contains a large number of water. On the opposite, the vacuum suction-type pneumatic conveyor is inefficient when the water content of the coal slime is low. So the design of the bucket chain convey-cleaning machine is designed to overcome those weaknesses [1]. The structure of the bucket chain convey-cleaning machine is shown in Fig. 1. Body frame and walking driver 9 are installed on the flatbed 10, sprocket driver 1 and oriented sprocket 4, 5, 6 are installed on the frame body, bucket chain 2 is hinged with chain 3 solidly, the pallets 7 solid joint with the frame body. When working, firstly start the motor, chain bucket convey-cleaning machine reach the working position with a tramcar, start the drive sprocket 1st motor, the chain roll in a clockwise way, walking driver drive the whole machine move forward, bucket chain move to the bottom, automatically loading the material, the machine tip when moving to bucket chain sprocket 6, pour the material into the tramcar"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001268_3-540-34319-9_8-Figure8.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001268_3-540-34319-9_8-Figure8.6-1.png",
        "caption": "Figure 8.6: Mecanum principle, turning clockwise (seen from below)",
        "texts": [
            " all four wheels being driven forward, we now have four vectors pointing forward that are added up and four vectors pointing sideways, two to the left and two to the right, that cancel each other out. Therefore, although the vehicle\u2019s chassis is subjected to additional perpendicular forces, the vehicle will simply drive straight forward. In Figure 8.5, right, assume wheels 1 and 4 are driven backward, and wheels 2 and 4 are driven forward. In this case, all forward/backward veloci- Omni-Directional Robots 8 ties cancel each other out, but the four vector components to the left add up and let the vehicle slide to the left. The third case is shown in Figure 8.6. No vector decomposition is necessary in this case to reveal the overall vehicle motion. It can be clearly seen that the robot motion will be a clockwise rotation about its center. Kinematics The following list shows the basic motions, driving forward, driving sideways, and turning on the spot, with their corresponding wheel directions (see Figure 8.7). \u2022 Driving forward: all four wheels forward \u2022 Driving backward: all four wheels backward \u2022 Sliding left: 1, 4: backward; 2, 3: forward \u2022 Sliding right: 1, 4: forward; 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003916_s0005117913080079-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003916_s0005117913080079-Figure1-1.png",
        "caption": "Fig. 1.",
        "texts": [
            " The present work is a continuation of [2] and differs in that here we take into account that the considered scheme has a support plate for each percussion piston whose levels can be controlled to tune the mechanism for a specific desired operation mode. We also give results of numerical experiments that have let us find different types of the mechanism\u2019s motion, including stochastic ones, and find values of parameters for which known period doubling bifurcations are realized. 2. PROBLEM SETTING The scheme of the considered mechanism is shown on Fig. 1a, where 1 is the mechanism\u2019s hull in which the bearings contain an axis 2 with a flywheel 3 and cranks 4 and 5 with radii r1, r2 that have a stationary phase shift \u03d5. The cranks have a joint connection to the piston rods 6 (of length li); percussion pistons 7 and 8 are, in turn, connected to them with a joint. Under the percussion pistons we have anvils 9, 10 with heights h1, h2 respectively. The hull 1 has a joint connection to the racks 11 that are rigidly connected to the anvil block 12. The rotation of the axis in the proposed scheme is transformed with a crank mechanism into a back-and-forth motion with respect to the anvil block",
            " The percussion pistons may alternate their blows to the corresponding anvils, and the vibroimpulsive impact created here is transmitted through the anvil block to the material being compacted. 1307 We denote by x the coordinate of the hull\u2019s center of mass M , by xpi (i = 1, 2), coordinates of the percussion piston masses mi, and compose expressions for kinetic V and potential \u03a0 energies V = 1 2 M x\u03072 + 1 2 m1x\u0307 2 p1 + 1 2 m2x\u0307 2 p2, \u03a0 = Mgx+m1gxp1 +m2gxp2, representing the Lagrange function as L = 1 2 M x\u03072 + 1 2 m1x\u0307 2 p1 + 1 2 m2x\u0307 2 p2 \u2212Mgx\u2212m1gxp1 \u2212m2gxp2. Using Lagrange equations of the second kind and the relation between hull coordinates and percussion pistons (see Fig. 1b), we get xp1 = x\u2212 s1 + r1 cos\u03c9t\u2212 \u221a l21 \u2212 r21 sin 2 \u03c9t, xp2 = x\u2212 s2 + r2 cos(\u03c9t\u2212 \u03d5)\u2212 \u221a l22 \u2212 r22 sin 2(\u03c9t\u2212 \u03d5); also, using obvious relations r1 l, r2 l and assuming li \u2248 l, i = 1, 2, we represent the mechanism\u2019s motion equations as d2x d\u03c42 \u2212 \u03bc\u03bb1 cos \u03c4 \u2212 \u03bc \u03b3\u03bb2 cos(\u03c4 \u2212 \u03d5) + p = 0 (x > f(\u03c4)), (1) dx d\u03c4 \u2223\u2223\u2223\u2223 + = \u2212R dx d\u03c4 \u2223\u2223\u2223\u2223\u2212 + (1 +R) df(\u03c4) d\u03c4 ( x = f(\u03c4), x\u0307\u2212 df d\u03c4 < 0 ) , (2) f(\u03c4) = max \u03c4 (f1(\u03c4), f2(\u03c4)), f1(\u03c4) = k1 + \u03b5\u2212 \u03bc cos \u03c4, f2(\u03c4) = k1 +\u0394k \u2212 \u03bc\u03b3 cos(\u03c4 \u2212 \u03d5), AUTOMATION AND REMOTE CONTROL Vol. 74 No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003687_detc2011-47626-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003687_detc2011-47626-Figure1-1.png",
        "caption": "Fig. 1 Cross section of bump-type foil bearings",
        "texts": [
            "org/about-asme/terms-of-use 3 Copyright \u00a9 2011 by ASME eR Reynolds number hoR , hiR Outside and insider radius of housing 0s Bump pitch ft , bt Foil thickness T Temperature 0T Ambient temperature hT Temperature of housing inT Temperature at inlet ,rec sucT T Temperature of recirculating flow and suction flow , ,u v w Air velocity component in x, y, z U Velocity of foil air s , h , b Thermal expansion coefficients  Foil deflection s , h , b Thermal expansion  , y Axial and angular coordinate  , m , n Constant  , 0 Viscosity of air  Bump foil Poisson\u2019s ratio cr x Detachment position of top foil , ,x y   Computational domain  F i  The ith interpolation coefficient of air fluidity  Density of air Since bump-type foil bearings possess higher load capacity than other types, GFBs with a bump-type structure are of increasing interest. Figure 1 shows a cross section of a typical bump-type foil bearing. The smooth top foil acting as the bearing surface is supported by a corrugated bump foil, which gives the shaft appropriate flexibility and frictional damping. Analysis of the foil structure in bump-type foil bearings shows that four factors, elasticity of bumps, friction forces, interaction forces and local deflection of the top foil, are essential for an accurate analytical model. In the mathematical model, one bump is replaced with a link-spring structure, which consists of two rigid links and one horizontally spaced spring, as shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003713_422-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003713_422-Figure1-1.png",
        "caption": "Figure 1. Schematic illustration of a slab shape specimen of bulk RE-123 oxide superconductor.",
        "texts": [
            " In this paper, a standard method of measurement of the irreversibility field for RE-123 bulk oxide superconductors is proposed. The measurement of the irreversibility field is tried, for two NEG-123 specimens, with the proposed standard measurement method. The target COV (coefficient of variation) is 5%. 0953-2048/05/120219+04$30.00 \u00a9 2005 IOP Publishing Ltd Printed in the UK S219 The specimen is of RE-123 bulk oxide superconductor prepared by the melt growth method. It was cut in a slab shape with the c-axis normal to the wide surface, as shown in figure 1. The external magnetic field is applied parallel to the c-axis. Then, the magnetic moment m is measured as a function of time, and the electric field E is obtained as E = \u2212 \u00b50G 2d(l + w) dm dt = \u2212\u00b50Gw2(3l \u2212 w) 24(l + w) dJ dt , (1) where G is a coefficient determined by the geometry of the specimen and the isotropic current density is assumed to be in the a\u2013b plane [5]. Since the relaxation of the magnetic moment due to the flux creep becomes smaller with time, E decreases to a sufficiently small value and E does not appreciably change during the period of measurement after some waiting time",
            " That is, it is sufficient to fix the size of the specimen and the waiting time after setting the external magnetic field for the standard measurement of the irreversibility field. The magnetic moment m is measured by a SQUID magnetometer or VSM. Since the unit of m is emu in many measurement systems, the unit must be converted to an SI unit as follows: m (A m2) = m (emu) \u00d7 10\u22123. (2) The critical current density is estimated from the magnetic hysteresis of the magnetic moment m using the Bean model [7]: Jc = 6 m w2 d(3l \u2212 w) (3) where d, l , w are defined in figure 1. 106 A m\u22122 is assumed as the criterion for the critical current density for the determination of the irreversibility field for RE-123 superconductors. For an exact estimation of the irreversibility field, five to ten measurement points seem to be enough within \u00b15% around the irreversibility field. Therefore, a temporary irreversibility field is first measured with a large step of the magnetic field. Then, accurate measurements are performed around the temporary irreversibility field with a small step of the magnetic field"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003913_amr.605-607.1212-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003913_amr.605-607.1212-Figure3-1.png",
        "caption": "Fig. 3 The calculation model Fig. 4 Condition 1 Fig. 5 Condition 2 The von Mises stress (Mpa) The von Mises stress (Mpa)",
        "texts": [
            "1 shows the general structure of the boom. The welded box beam is adopted by the down chord as a small car track. The root of the down boom is in the form of bifurcation structure so that it can let the lower chord and the turntable connected directly, as shown in Fig. 2. And the cable is used with the structure form of double-point and four-lasso which connects with the lord limb of the top boom. The article has taken the root\u2019s structure of the down boom as calculation model. The calculation model is showed in Fig. 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. (ID: 141.211.4.224, University of Michigan Library, Media Union Library, Ann Arbor, USA-08/07/15,14:54:38) Conditions. The loads of the down chord are only considered its own weight and the axial force of the oblique truss, and the axial force of the oblique truss are based on the following four conditions. The first one, the cargo boom hoists the maximum load when being in the minimum range",
            " The second one, the cargo boom hoists the maximum load in the short lasso place. The third one, the cargo boom hoists the maximum load based on the lifting characteristics when the lifting hook is in the position between the short and long lasso. The forth one, the cargo boom hoists the maximum rated load when working in the maximum range. The von Mises stress (Mpa) The von Mises stress (Mpa) 1\uff0c2\uff0c3and4 in the Table 1-6 represent the position of the rigid connection of the oblique truss and the down chord, referring to the Fig. 3. to the surrounding, referring to the Fig. 3. Table 2 The results of the web position von Mises stress [Mpa] Condition 1 Condition 2 Condition 3 Condition 4 1 150 101 69 218 146 101 247 166 114 329 22 1 152 2 137 91 62 201 134 91 232 155 105 315 21 0 143 3 142 108 65 60 41 27 65 41 29 82 59 43 4 66 53 41 60 42 29 72 45 31 97 60 42 Notice: Each condition\u2019s von Mises stress is derived from the top-down of rigid connection point of the wed. The conclusions can be drawn from Fig. 4-7 and Table 1- 3 as follows: From the Table 1, due to the influence of the oblique truss's axial force, the larger local stress nearby the joint of the oblique truss and the down chord will be produced and the local stress of the top flange at the center of connection position will reach the maximum, and then reducing gradually to the surrounding"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002114_978-1-4471-4141-9_19-Figure19.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002114_978-1-4471-4141-9_19-Figure19.3-1.png",
        "caption": "Fig. 19.3 The finite element analysis model",
        "texts": [
            " When we fit the deformation function under load with Fourier function, there are several factors to judge the fitting quality including SSE (Sum of squares due to error), RMSE (Root mean squared error), Adjusted R-square (degree-of-freedom adjusted coefficient of determination), R-square (Coefficient of determination), in which if the first two are close to 0, and the latter two approach to 1, indicating the fitting effect is pretty good. On the contrary, the fitting effect is bad. To acquire more reasonable conjugate tooth profile, the calculation procedure of the deformation function should be a multi-iterative process. The process flow diagram is shown in Fig. 19.2: 1. Finite element analysis model: the finite element analysis model of Abaqus is established shown in Fig. 19.3. The material parameters of each component are listed in Table 19.1. The model is modelled with C3D8R element; the solving process is divided into two steps. The first step is simulating the process that the wave generator is installed into the flexspline, forcing the flexspline to produce elastic deformation and prestress. The outer ring of the rigid circular spline is under full constraints, with implicit algorithm to calculate this step; The second step based on the first step add an rated speed n to the wave generator, and torque load T is applied on outer ring of the output circular spline"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002407_amr.189-193.1409-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002407_amr.189-193.1409-Figure3-1.png",
        "caption": "Fig. 3. The SPS limb which afford no cconstraint",
        "texts": [
            " Since the limb constraint system consist one constraint force 1f , the twists of limb can be obtained: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= , 1 4 4 4[0,0,0; , , ]P d e f= , 2 5 5 5[0,0,0; , , ]P d e f= The new revolute joints 4 5,R R can be obtained by the linear combination of revolute joint and prismatic joint, so the twist system of limb could become: 1 1 1 1[ , , ;0,0,0]R a b c= , 2 2 2 2[ , , ;0,0,0]R a b c= , 3 3 3 3[ , , ;0,0,0]R a b c= 4 1 1 1 4 4 4[ , , ; , , ]R a b c d e f= , 5 2 2 2 5 5 5[ , , ; , , ]R a b c d e f= Therefore, the UPU limb can be obtained to afford the constraint force 1f , as shown in Fig. 2. In the similar method, the SPS limb can be obtained to afford the constraint no force, as shown in Fig. 3. Finally, The symmetrical decoupled SPM (Fig. 4-a) and non-symmetrical decoupled SPM(Fig. 4-b) can be obtained by assembling the limbs synthesized perpendicularly. Since each UPU limb affords one constraints f1, the 3 UPU\u2212 parallel mechanism (Fig. 4-a) and 2SPS RU\u2212 parallel mechanism (Fig. 4-b) is non-over-constrained. The principle of decoupled motion of UPU RU SPS\u2212 \u2212 SPM are as following: when the active prismatic joint of UPU chain works, the moving-platform of spherical parallel mechanism rotates around axis X (actuated by the motor 1), while the actuators of two kinematic chains holds as the kinematic chain has the kinematic joint along axis X"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002191_978-94-007-1415-1_3-Figure3.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002191_978-94-007-1415-1_3-Figure3.12-1.png",
        "caption": "Fig. 3.12 Dual rod pendulum. (a) In rest position. (b) Deflected",
        "texts": [
            " In consequence, the time of swing, T , of the trapezium pendulum decreases relative to that of a simple rod pendulum. In the special case where l2 D l3 the pilot crank and rocker of a four bar linkage are parallel, and a trapezium pendulum becomes a dual rod pendulum. A dual rod pendulum consists of two rigid massless rods, length l , suspended from frictionless pivots connected to a horizontal rigid massless rod with a point mass, m, at its centre. In the rest position the rigid massless rods, length l , are vertical, as shown in Fig. 3.12a. The motion of the point mass, m, is identical to that of a simple rod pendulum (Sect. 2.3) except that, because of interference between the links, the pendulum angle, , (Fig. 3.12b) cannot exceed 90\u0131. For small amplitudes the time of swing, T , is given approximately by Eq. 2.13 where g is the acceleration due to gravity. A double string pendulum (Fig. 3.13) consists of two simple string pendulums (Fig. 1.4b) arranged in series. The upper inextensible massless string, length l1, is clamped at its upper end, and has a point mass, m1, at its lower end. The lower inextensible massless string, length l2, is attached at its upper end to the point mass m1, and has a point mass, m2, at its lower end"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.33-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.33-1.png",
        "caption": "Figure 12.33 A rigid body B in a general motion.",
        "texts": [
            " The constant m is the mass of the body and GI is the mass moment matrix of the body as calculated in G. We attach a body coordinate frame B (oxyz) to the rigid body at C and express the equations of motion in B: BF = m B GaB + m B G\u03c9B \u00d7 BvB (12.524) BM = BI B G\u03c9\u0307B + B G\u03c9B \u00d7 ( BI B G\u03c9B ) (12.525) Although the B- and G-expressions of equations of motion can equivalently describe the motion of a rigid body, the B-expression of the Euler equation has great advantage by having a constant mass moment matrix GI . Proof : Figure 12.33 illustrates a rigid body B that has a general motion in the global coordinate frame G. We attach a body frame B at a point o. Application of the Euler theorem about the rigid-body motion with a fixed point allows us to analyze the general motion of a rigid body by considering the motion of a point o plus the motion of the body about o. To show this, let us determine the kinetic energy of the body. Point o is an arbitrary fixed point in B, and point C is the mass center of B. The velocity of a mass particle dm at P is GvP = Gr\u0307P = Gd\u0307B + G\u03c9B \u00d7 (GrP \u2212 GdB ) = Gd\u0307B + G\u03c9B \u00d7 G B rP (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002522_amr.346.332-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002522_amr.346.332-Figure2-1.png",
        "caption": "Fig. 2 A planar aerostatic bearing with single restrictor",
        "texts": [
            " 5)Calculate the stiffness matrix of the entire aerostatic bearing according to (7). The proposed formulation can be used to calculate the stiffness characteristics of aerostatic bearings with various restrictor designs and for guideways with various shapes of gas film such as linear, planar, spherical and cylindrical surfaces. The resulting stiffness matrix contains both the normal stiffness and the roll stiffness. Roll vibration analysis of the system can thus be figured out. 3.Numerical examples The planar aerostatic bearing as shown in Fig. 2 is taken as case study to illustrate the proposed method. The parameters of the system are listed as bellows: 1 40 mmr = , 2 1.5 mmr = , 0.2 mmd = , 1 0.4 mmh = , 2 0.1mmh = , 1 3.0 barp = . The atmospheric pressure 1.013 barap = . The relationship between the pressure distribution p and the air gap h solved by using the CFD software Fluent is shown in Fig. 3. The pressure p at random position is irrelative to \u03b8 but depends on the air gap h and the radius r at the point. The derivative of p with respect to h is obtained by using finite difference method and the results are shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003416_amm.271-272.1032-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003416_amm.271-272.1032-Figure2-1.png",
        "caption": "Fig. 2 Cross-section of the journal bearing.",
        "texts": [
            " The purpose of the present paper is to explain the mechanism of vibrations and reveal its dynamic characteristics of the flexible rotor system with a misaligned journal bearing, thereby, for the above symmetrical rotor system, then its motion equations can be expressed )( )( )sin()()( )cos()()( )( )( 2333 2333 2 32122 2 32122 2111 2111 yykFym mgxxkFxm tMayykyykyM tMaMgxxkxxkxM yykFym mgxxkFxm y x y x \u2212\u2212= +\u2212\u2212= \u2126\u2126+\u2212\u2212\u2212\u2212= \u2126\u2126++\u2212\u2212\u2212\u2212= \u2212\u2212= +\u2212\u2212= (1) where F1x, F1y, F3x, F3y are the oil film forces of journal bearings in x and y directions, and other parameters are presented in Fig. 1. In this work, the long bearing theory is introduced, thus, the Reynolds equation for the oil film pressure, p, in Figure 2 yields \u03b8 \u03b8 \u03d5 \u03b8\u00b5\u03b8 cos)2( 2 1 12 1 3 2 e hph R + \u2202 \u2202 \u2212\u2126=      \u2202 \u2202 \u2202 \u2202 (2) where )cos1(cos \u03b8\u03b5\u03b8 +=+= cech . Therefore, the oil film forces in radial and tangent directions are gotten by integrating over the area of the journal sleeve under the boundary condition 00 == == \u03c0\u03b8\u03b8 pp and half-Sommerfeld condition 2 2 1 2 0 2 2 2 3 4 0 2 ( ) cos 6 ( 2 ) 2 ( ) sin 6 ( 2 ) 2 B j j Bjr j j j j j B j j Bjt j j j j j d dR F p R dzd BR E E c dt dt d dR F p R dzd BR E E c dt dt \u03c0 \u03c0 \u03d5 \u03b5 \u03b8 \u03b8 \u03b8 \u00b5 \u03d5 \u03b5 \u03b8 \u03b8 \u03b8 \u00b5 \u2212 \u2212  = = \u2126 \u2212 +         = = \u2126 \u2212 +        \u222b \u222b \u222b \u222b j=1,3 (3) where 2 1 2 2 2 (1 )(2 ) j j j j E \u03b5 \u03b5 \u03b5 = \u2212 + \uff0c 2 2 3/ 2 2 1 8 (1 ) 2 (2 ) j j j E \u03c0 \u03b5 \u03c0 \u03b5   = \u2212  \u2212 +   \uff0c 3 2 1/ 2 2(1 ) (2 ) j j j j E \u03c0\u03b5 \u03b5 \u03b5 = \u2212 + \uff0c 4 2 2 2 (1 )(2 ) j j j j E \u03b5 \u03b5 \u03b5 = \u2212 + "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002982_amr.317-319.281-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002982_amr.317-319.281-Figure3-1.png",
        "caption": "Fig. 3. Structure of the flexspline",
        "texts": [
            "2, circle 1 stands for the nominal profile of the flexspline contacting the wave generator without any gap or contact deformation, circle 2 stands for the wave generator profile with a gap \u03b4uniformly distributed along the whole profile, and circle3 (dashed) stands for the flexspline profile under loads. Under certain loads, some areas of the flexspline are closer to the wave generator and thus deformed, while other areas do not contact the wave generator and yield distortion. When the wave generator rotates in clockwise direction, the points on arc AB\u2032and A B\u2032 cannot deform freely in the radial direction. The harmonic gear drive is mainly composed of flexspline, output shaft, wave generator and circular spline. The dimensions of the flexspline are depicted in Fig.3. The1/4 or 1/2 of the flexspline finite element model is often chosen for simulation under no-load condition. Under load conditions, however, the force between the flexspline and the wave generate is not axial symmetry because of their bilateral symmetry. Therefore, it is not suitable to choose 1/4 or 1/2 of the flexspline model to investigate the transient dynamic problem. Instead, the whole model of the flexspline is to be applied in this work. Modeling method To obtain more accurate results, we use 8-node hexahedral element to model flexspline, circular spline and wave generator",
            " In the harmonic gear drive, there are some contact surfaces between wave generator and flexspline, and between circular spline and flexspline. Since solving the problem is computationally intensive, we define two contact regions in Fig.4 to reduce the contact search time and neglect the effects of wave generator bearing balls in the model to increase the compute speed. The transient dynamic analysis is carried out using the model shown in Fig. 4 under the 1000rpm rotational speed and 25N\u00b7m load, on the basis of large displacement and large strain theory. In the Fig. 3(b), sections of flexspline start from flexspline cup rim along the axial direction, having the lengths of 0.04, 0.16, 0.24, 0.27, 0.28 and 0.3 times of flexspline diameter, respectively. The stress variation curves on these sections are shown in Fig.5. Tab. 1 Material parameters of harmonic gear drive Components of harmonic gear drive Flexspline Circular spline, wave generator and output shaft Material type 30CrMnSi 45# Steel Tensile modulus [GPa] 204 210 Poisson's ratio \u00b5 0.29 0.3 Density \u03c1 [g\u00b7cm -3 ] 6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002027_978-3-319-00777-9_25-Figure25.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002027_978-3-319-00777-9_25-Figure25.2-1.png",
        "caption": "Fig. 25.2 DCB test specimen",
        "texts": [
            " At least five samples were tested for each material type. Fracture toughness was determined using the Double-Cantilever Beam (DCB) test as outlined in ASTM D5528. To manufacture samples for this test, composite panels were fabricated following the normal procedure above while inserting a Teflon strip in the mid-plane of the laminate stack. This had the effect of creating an initial crack in the material as required by the standard. Test samples were then prepared from these panels as shown in Fig. 25.2, having a width of 20\u201325 mm and a length of 125 mm. These samples were loaded into an INSTRON testing machine and loaded at a crosshead speed of 0.5 mm/min while the delamination length was monitored using a digital microscope. Data reduction was performed using a robust numerical method to best capture the strain energy involved in crack growth. At least five samples were tested for each material type. To perform moisture absorption tests, 100 by 300 test samples were cut from the composite panels following ASTM D570"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.34-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.34-1.png",
        "caption": "Fig. 2.34 4PaPPR-type fully-parallel PMs with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology Pa||P\\P||R (a) and Pa\\P\\kP\\kR (b)",
        "texts": [
            "30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20. 4PaPRP (Fig. 2.35b) Pa||P\\R||P (Fig. 2.21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001669_978-3-642-33832-8_1-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001669_978-3-642-33832-8_1-Figure3-1.png",
        "caption": "Fig. 3 Wear rate with different accelerations and speeds",
        "texts": [
            " For a hard self-excited vibration, only if the initial exciting force is large enough to push the phase trajectory to pass the unstable limit cycle S2 and to achieve the stable limit cycle S1; otherwise, it will converge to the stable point O2:; which is shown in Fig. 2b. By analyzing the polygonal wear of tire, it is concluded that the tread\u2019s vibration is a typical hard self-excited vibration. Lupker [11] pointed out that the average wear rate of rubber would increase nonlinearly with the augment of vibration acceleration. As shown in Fig. 3, with the increase of vehicle speed, the normalized wear rate gets higher gradually, and then become smaller after the vehicle speed of 100 km/h. In another word, for a tire\u2019s self-excited vibration at the same frequency, different vibration accelerations can generate different vibration amplitudes, which cause different normalized wear rates. Therefore, it should be focused to control the tread\u2019s vibration amplitude or to move the vehicle speed that can cause hard self-excited vibration of tire out of the normal speed range"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure18-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure18-1.png",
        "caption": "Fig. 18 Typical End Mill installation for Tip chamfering\u2014Spur and straight bevel gears",
        "texts": [
            " 15 \u201cAC\u201d type CnC machine limits for Toe/Heel chamfering\u2014large pitch cone angle bevel gear The same comments apply when the pitch cone angle of a straight bevel gear is large, as is shown in Fig. 17. Tip chamfering can also be performed using the EM tool\u2019s tip which, in most cases, is a better approach since the turn table tilt is less and there are no risks of tool interference. However, the tool diameter becomes a limiting factor to avoid damaging the opposite tooth flank. When chamfering the Tip edges, the EM toolmust be tangent to the axial direction of the tip and angled to provide the desired chamfer (Fig. 18). Again, five unit vectors (left, Fig. 18) are required at any point along the tooth Tip edge: N : the local normal vector, 164 C. Gosselin \u2212\u2192 Vo: N X T ,\u2212\u2212\u2212\u2192 Trans: the axis about which the tool vector is pivoted,\u2212\u2212\u2192 Tool: the tool vector. \u2212\u2192 Vo is obtained from the cross product of N and T . Vector \u2212\u2212\u2212\u2192 Trans is obtained by pivoting T about \u2212\u2192 Vo of \u03c0 2 \u2013 the local pressure angle; pivoting T about \u2212\u2192 Vo to bisect the Tip edge (right, Fig. 18), and about \u2212\u2212\u2212\u2192 Trans by the required pivot angle to ensure that the tool is perpendicular to the axial direction of the Tip edge yields vector \u2212\u2212\u2192 Tool (Fig. 19). In the above, Toe and Heel chamfering are performed with neither rotation of the work piece\u2014apart from indexing\u2014nor change in turn table tilt which improves the predictability of the movements, especially when the work piece is large and heavy. This also implies that 4 and 4+1 axes machines can be used. By contrast, Tip chamfering of bevel gear teeth requires a continuous reorientation of the tool axis in reference to the work piece axis and, therefore, a 5 Axis CnC machine is required"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002693_amr.605-607.175-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002693_amr.605-607.175-Figure1-1.png",
        "caption": "Fig. 1 Process of identification, orientation and delivering",
        "texts": [
            "199, Purdue University Libraries, West Lafayette, USA-09/07/15,05:33:03) and orientation device which is fit in the middle of the discharging chute. It\u2019s used to receive the sleeve from the discharging chute, and then the sleeve is delivered to the next discharging chute after identification and orientation. In the sleeve assembly process, the crevice direction of the sleeves on the platform are random distribution, the crevice orientation is realized by rotating recognition after which the sleeve is delivered to the next feeding tube. The device includes two function parts (shown in Fig. 1): the rolling the sleeve component and the delivering the sleeve to the next feeding tube. The device is comprised by a collet, a drive sprocket and two driven sprockets. The drive sprocket is driven by a stepping motor and the two driven sprockets by another stepping motor through a lead crew to drive. While the crevice direction of the sleeve on the platform is random, the driven sprockets approach to the drive one (shown in Fig. 1a) until the sleeve is clamped (shown in Fig. 1b) when the drive sprocket begin to rotate driving the sleeve rolling under the friction action. The sleeve crevice is recognized by a photoelectric sensor, when it is found, the drive sprocket stops rolling while the driven one begin to leave and the sleeve also leave in the collet (shown in Fig. 1c), therefore the sleeve will fall in to the next feeding tube. Drive and driven sprocket design. Fig. 2 shows the force analysis of the sleeve in the process of champing. When the sleeve is champing by the driven sprockets, the drive sprocket rotates driving the sleeve rolling under the friction action, at this time the sleeve is applied f1M , 1F from the drive sprocket, 2F , 3F , f2M , f3M from the driven sprockets and 1M from the sliding friction between the sleeve and the platform. Formulation 1 is obtained on the basis of coplanar system of concurrent forces to be in equilibrium and the force of rolling friction equation",
            " 1 2 3 1 2 3 1 cos cos 0 0f f f F F F M M M M \u03b8 \u03b8\u2212 \u2212 =  \u2212 \u2212 \u2212 \u2265 (1) In which, 1 1 1fM F\u03b4= , 2 3 2 2f fM M F\u03b4= = , 1M Gh\u00b5= , 1\u03b4 is the friction coefficient between the drive sprocket and the sleeve, 2\u03b4 is the friction coefficient between the driven sprockets and the sleeve, \u00b5 is the sliding friction coefficient between the platform and the sleeve, G is the gravity of the sleeve and h is the arm of couple between the sliding force and the balanced surface. (2) can be deduced from (1). The sleeve will roll without slip in the condition as (2) shows. 2 1 2 cos 2 Gh F \u03b4 \u00b5 \u03b8 \u03b4 \u2265 + (2) Collet design. The collet parameter analysis. The collet deforms two times in the process of sleeve crevice identification and orientation as Fig. 1a and Fig. 1c shown. The collet is a symmetrical structure, so the stress and deformation on either side are the same, which is why the paper chooses one of the sides to analyze. Fig. 3a shows the collet deformation when the sleeve is clamping, fig. 3b shows the deformation when the collet is resetting, at which time the sleeve is being delivered to the next feeding tube. There are bending moment and shearing force on the transverse of collet at the same time. It is an assumption that the collet is a bar, on which the shearing force influence can be ignored"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001932_9783527680436.ch4-Figure4.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001932_9783527680436.ch4-Figure4.2-1.png",
        "caption": "Figure 4.2 Principle sketch of a microbial fuel cell and its major reactions.",
        "texts": [
            " In fact, the main driving force is the development of a platform technology for different sustainable applications, based on microbial bioelectrocatalysis, often referred to as \u201cmicrobial bioelectrochemical systems\u201d (BES) [5]. 4.2 Microbial Bioelectrochemical Systems (BESs) 4.2.1 Probably more than 90% of current BES research is dedicated to microbial fuel cells (MFCs), the archetype microbial bioelectrochemical system. Although microbial fuel cells were long seen as a scientific peculiarity the first systematic MFC studies were performed in the framework of the NASA space program in the 1960s [3,4,6], later followed by Wilkinson and coworkers aiming at their application in robotics [7\u20139]. Figure 4.2 depicts the working principle of a microbial fuel cell. At the anode the oxidation of substrates, which may range from acetate to complex wastewater Electrocatalysis: Theoretical Foundations and Model Experiments, First Edition. Edited by Richard C. Alkire, Ludwig A. Kibler, Dieter M. Kolb, and Jacek Lipkowski. 2013 Wiley-VCH Verlag GmbH & Co. KGaA. Published 2013 by Wiley-VCH Verlag GmbH & Co. KGaA. constituents (see [10] for an overview), takes place, whereas the oxygen reduction reaction (ORR) represents the common cathode reaction [11]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003336_20110828-6-it-1002.03057-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003336_20110828-6-it-1002.03057-Figure3-1.png",
        "caption": "Fig. 3. Three rotating inertias interacting by friction (double dry clutch).",
        "texts": [
            " This result is quite intuitive since the static friction opposes the motion and, therefore, counteracts all the external torques in order to guarantee that the relative velocity remains zero. When z2 = 0, the system is completely described by state variable z1 and by the equation (8); as shown in Fig. 2(b) it behaves like a unique inertia J1 + J2, subject to the resultant of all the external torques. Note that, by substituting \u03c41,2 = J2F1\u2212J1F2 J1+J2 in both the equations of the original system (5) we just obtain (8). Given the system composed by three rotating masses shown in Fig. 3, that are subject to external torques and friction between the contacting masses, as described by the system of differential equations    J1\u03c9\u03071 = F1 \u2212 \u03c412(\u03c91 \u2212 \u03c92) J2\u03c9\u03072 = F2 + \u03c412(\u03c91 \u2212 \u03c92)\u2212 \u03c423(\u03c92 \u2212 \u03c93) J3\u03c9\u03073 = F3 + \u03c423(\u03c92 \u2212 \u03c93) (10) one may decouple the main dynamics that does not depends on the internal frictional torques from the other relative dynamics. By considering the state vector z =   J1\u03c91+J2\u03c92+J3\u03c93 JTot \u03c91 \u2212 \u03c92 \u03c92 \u2212 \u03c93   related to the vector of the velocities by \u2126 = Tz, with T =     1 J2+J3 JTot J3 JTot 1 \u2212 J1 JTot J3 JTot 1 \u2212 J1 JTot \u2212J1+J2 JTot     where JTot = J1 + J2 + J3, the system is translated into the form JT z\u0307 = F T \u2212DT \u03c4 (D T Tz) (11) with JT =     JTot 0 0 0 J1(J2+J3) JTot J1J3 JTot 0 J1J3 JTot (J1+J2)J3 JTot     F T =    FT 1 FT 2 FT 3    =     F1 + F2 + F3 \u2212J1(F2+F3)+(J2+J3)F1 JTot J3(F1+F2)\u2212(J1+J2)F3 JTot     , DT =    0 0 1 0 0 1    "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003695_dscc2012-movic2012-8715-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003695_dscc2012-movic2012-8715-Figure1-1.png",
        "caption": "Figure 1. Modeling of the elastic beam: coordinate system - used variables [7]",
        "texts": [
            " Up to now no direct comparison between this previously published approaches [1, 2] and [7] has been made. With this contribution the comparison is made for the first time. Simulations of the flexible beam applying the different angular velocities show the effect of the geometric stiffness on the deformation of the beam with and without the nonlinear terms, the results are compared with other results presented in several contributions [1, 2]. The used coordinate system and variables to model the considered elastic beam are shown in Fig.1. It is attached with a rigid base, which is driven by an angular velocity about the z-axis. The used coordinate systems of the beam are illustrated in Fig.1. Frame O\u2212 X IY IZI is the inertial reference frame. The moveable joint coordinate system G\u2212XGY GZG is a beam-fixed frame with its origin located at the centroid of the cross-section. The frame E \u2032\u2212XEY EZE is the reference frame defined for the crosssection to describe the elastic variables ux,uy,uz in the moveable joint coordinate system. The frame P\u2212XPY PZP is defined for the position and orientation of the considered infinite section. The equation of motion for the rotationg elastic beam are obtained by using the principle of virtual work \u03b4Ve = \u03b4Wa +\u03b4Wm, (1) where Ve, Wa, and Wm denote the virtual potential energy, the virtual work of the inertia forces, and the external forces, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003248_transducers.2013.6626747-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003248_transducers.2013.6626747-Figure3-1.png",
        "caption": "Figure 3: Our proposed Redox sensor functioned current amplifying. The chip is integrated with WE, CE, and bipolar transistor. The WE current signal is directly incoming to base of the transistor.",
        "texts": [
            " The applied potential from CE was shifted about 0.7 V for the voltage drop of bipolar transistor, as expected. The output current was amplified 60 times. The current amplifications in redox measurements were confirmed to observe the large amplification. Figure 2: Comparative measurement of effecting bipolar transistor using 2 mM of potassium ferricyanide solution. The current signal was increased by current amplifier. The voltage shift was 0.7 V for forward voltage of bipolar transistor, as expected. Figure 3 shows a schematic cross section of proposed Amplified Redox Sensor. The proposed redox sensor integrated electrodes and bipolar transistor on a chip was fabricated using LSI technology. The developed sensor is fabricated with an npn bipolar transistor. The bipolar transistor was fabricated on Si substrate. WE is connected to base in bipolar transistor, and CE were formed at the same chip. This structure has strong points which are small element area in spite of integrating bipolar transistor, about 100 times of large current amplifying, and lower effect of inpouring disturbance noise to output line for capable of large signal reading out"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002709_arso.2013.6705503-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002709_arso.2013.6705503-Figure6-1.png",
        "caption": "Fig. 6. Robots for modeling examples (left: TITAN XIII,right: GRYPHON)",
        "texts": [
            " Despite this, models can be further detailed on the fly by other modelers that may want to state or know more specific information about the system through the definition and merging of Roles, and they can use the same query mechanism independently from the detail level or abstractions used. Additionally, when using RDF files, it becomes compatible with the Semantic Web. In order to validate the proposed framework (and its implementation), it was applied into three example applications. In this example, two independent robot systems available at our laboratory were modeled using the FIERRo. They were the TITAN XIII and GRYPHON robot systems shown in Fig. 6. Each of the robots has different purposes and architectures (software and hardware). TITAN XIII is a quadruped robot intended for gait experimentation on irregular terrains [15]. GRYPHON is a robotic arm equipped with various sensors and mounted on a buggy intended for humanitarian demining operations [16]. The resulting diagrams for each of the systems are shown in Fig. 7 and Fig. 8. The framework provides the freedom of choosing the level of detail, in this example the TITAN XIII system model has a greater level of detail than the GRYPHON system model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003914_s1068798x13030052-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003914_s1068798x13030052-Figure1-1.png",
        "caption": "Fig. 1. Plastic deformation of the tool\u2019s cutting section in the case of a T15K6 alloy tool and a steel 45 blank (cutting time v = 3 m/s; supply s = 0.3 mm/turn; cutting depth 2 mm): h\u03b3, l\u03b3, modified height and length of the tool\u2019s front surface; h\u03b1, l\u03b1, modified height and length of the tool\u2019s rear surface; t, time of cutter operation, s.",
        "texts": [
            " When nor mal loads exceed the permissible flexural stress at the tool\u2019s front surface, the resulting brittle macroflaws break down the basic part of the cutting plate beyond the limits of contact at the front face. As a rule, the brit tle macroflaws appear in the initial period of cutting. Research on the plastic deformation of the tool\u2019s cutting section shows that cutting is associated not only with tool wear at the rear surface but also with lowering of the front surface and buckling of the rear surface (Fig. 1) [2]. Wear of the tool\u2019s rear surface and deformation of the cutting edge occur in parallel over time. In Fig. 1, we choose the same scale for the modified height of the cutter surface perpendicular to the initial position of the tool\u2019s front surface and the cutter wear at the rear surface (the modified height at the rear surface) and a different scale for the modified length of the tool\u2019s front surface. Likewise, one scale is adopted for the buckling (the modified height at the rear surface) perpendicular to the rear surface, and another for the modified length of the rear surface. Thus, the dimensions of the drop in the tool\u2019s front surface and the buckling of the tool\u2019s Optimizing Tool Life in Numerically Controlled Machine Tools E"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.27-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.27-1.png",
        "caption": "Figure 10.27 The magnetic field distribution of PM machines at no-load conditions (the stator current is zero): (a) a four-pole SPM motor; (b) an eight-pole symmetrical IPM motor; and (c) a four-pole unsymmetrical IPM configuration",
        "texts": [
            " It is worth pointing out that although there is a reluctance torque component, it does not necessarily mean an IPM motor will have a higher torque rating than a SPM motor for the same size and same amount of magnetic material used. This is because, in IPM motors, in order to keep the integrity of the rotor laminations, there are so-called \u201cmagnetic bridges\u201d that will have leakage magnetic flux. So for the same amount of magnet material used, a SPM motor will always have higher total flux. There are many different configurations for IPM motors as shown in Figure 10.26. The no-load magnetic field of PM machines is shown in Figure 10.27. When the rotor is driven by an external source (such as an engine), the rotating magnetic field will generate three-phase voltage in the three-phase windings. This is the generator mode operation of the PM machine. When operated as a motor, the three-phase windings, similar to those of an induction motor, are supplied with either a trapezoidal form of current (brushless DC) or sinusoidal current (synchronous AC). These currents generate a magnetic field that is rotating at the same speed as the rotor, or synchronous speed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.7-1.png",
        "caption": "Figure 12.7 A two-link manipulator.",
        "texts": [
            "130), which confirms that the dynamic characteristics of rigid bodies in a principal frame are determined by the eigenvalues and eigenvectors of [I ], not by the actual geometric shape of the body. Transforming 2M to B1, we have 1M = 1R2 2M = \u23a1 \u23a2 \u23a3 1 2 (I1 \u2212 I2) 2 sin 2\u03b82 \u03c9 (I2 \u2212 (I1 \u2212 I2) cos 2\u03b82) \u2212 (I1 \u2212 I2) \u03c9 sin 2\u03b82 \u23a4 \u23a5 \u23a6 (12.141) If is very small, then 1Mx is almost zero. However, the required torque to turn the shaft increases rapidly with increasing . Example 724 Angular Momentum of a Two-Link Manipulator A two-link manipulator is shown in Figure 12.7. Link A rotates with angular velocity \u03d5\u0307 about the z-axis of its local coordinate frame. Link B is attached to link A and has angular velocity \u03c8\u0307 with respect to A about the xA-axis. Let us attach coordinate frames A and B to links A and B such as shown in the figure. We assume that A and G were coincident at \u03d5 = 0; therefore, the transformation matrix between A and G is GRA = \u23a1 \u23a3 cos \u03d5(t) \u2212 sin \u03d5(t) 0 sin \u03d5(t) cos \u03d5(t) 0 0 0 1 \u23a4 \u23a6 (12.142) The frame B is related to A by Euler angles \u03d5 = 90 deg, \u03b8 = 90 deg, and \u03c8 = \u03c8 ; hence, ARB = \u23a1 \u23a3 c\u03c0c\u03c8 \u2212 c\u03c0s\u03c0s\u03c8 \u2212c\u03c0s\u03c8 \u2212 c\u03c0c\u03c8s\u03c0 s\u03c0s\u03c0 c\u03c8s\u03c0 + c\u03c0c\u03c0s\u03c8 \u2212s\u03c0s\u03c8 + c\u03c0c\u03c0c\u03c8 \u2212c\u03c0s\u03c0 s\u03c0s\u03c8 s\u03c0c\u03c8 c\u03c0 \u23a4 \u23a6 \u23a1 \u23a3 \u2212 cos \u03c8 sin \u03c8 0 sin \u03c8 cos \u03c8 0 0 0 \u22121 \u23a4 \u23a6 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003511_amm.198-199.171-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003511_amm.198-199.171-Figure5-1.png",
        "caption": "Fig. 5 Schematic Diagram of the Mechanism",
        "texts": [
            " When the collection of balls is put into the serving mechanism, it can provide the tennis-serving practice. In the experimental testing, the design of the tennis vehicle can achieve good results of auxiliary tennis practice in the tennis court. Serving Mechanism. The tennis vehicle proposed by this paper is integrated technical products which makes the electronic, mechanical and computer technology as a whole. The serving mechanism are used the principle of a single spinning wheel to work, which is shown in Fig. 5(a). It is mainly composed of a friction wheel and a fixed raceway, and the gap of the wheel and the raceway is less than the diameter of the tennis balls. The wheel is driven by a direct current motor. Tennis balls are applied the pressure and friction force in the raceway by the wheel. When the balls get out of the raceway, they can obtain the initial velocity and launch out. This principle is simple and practical, which can ensure the stability of the launched balls and well control the frequency and speed of the tennis. Picking Mechanism. The picking mechanism picks up balls into the collection box by friction. Schematic diagram of the mechanical structure of the picking mechanism is shown in Fig. 5(b). 1-Drum, 2-Shell are driven to scroll up by the rolling drum, which can achieve the purpose of collecting balls. The mechanism has the structural design gathering balls, and the driving wheels of the picking mechanism are directly merged with the vehicle in order to reduce the overall size. At the same time, it transmits driving force to the drum to realize picking up balls. Control system uses MCS-51 MCU [6] as the core microprocessor, and peripheral circuit includes motor drive circuits, a remote control circuit and a detection circuit for the balls"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002487_s11044-013-9384-5-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002487_s11044-013-9384-5-Figure1-1.png",
        "caption": "Fig. 1 The side view of the mechanical model with its parameters. As full left\u2013right symmetry is assumed, the two oars are modelled as one unit",
        "texts": [
            " The boat was given one translational dof, such that only movement in the lengthwise horizontal direction was considered as vertical oscillations have been shown to have minimal impact on the system dynamics, [22], and transversal horizontal movements were seen as minor disturbances in the symmetric movement. The foot segment was assumed to be fixed at the footstretcher, i.e. to have the same velocity as the boat (no rotation). The shank, thigh, and trunk were assumed to give in-plane motions through the rotational dofs q2, . . . , q4, Fig. 1, neglecting the more detailed movement pattern in the spine of the rower, [23, 24]. The arm was allowed to have out-of-plane motion through two rotational dofs at the shoulder (q5, q6), and one rotation at the elbow (q7). The two oars, modelled as one unit, were given two rotational dofs, (q8, q9). Five control forces, i.e. torques were used to drive the model and act in the physical joints: ankle, knee, hip, shoulder (flexion/extension), and elbow. Four coupling forces were used to represent the conditions at the slider (c4), and the three force components at the oar handle (c7, ",
            " The forces were introduced as additional terms in the weak form of the equilibrium equations as \u222b T 0 ( c7(t)\u03b4 ( Xoar handle(q) \u2212 Xhand(q) ) + c8(t)\u03b4 ( Yoar handle(q) \u2212 Yhand(q) ) + c9(t)\u03b4 ( Zoar handle(q) \u2212 Zhand(q) )) dt (13) and the discrete constraints like: Xoar handle(q) = Xhand(q), (14) Yoar handle(q) = Yhand(q), (15) Zoar handle(q) = Zhand(q). (16) At the seat, contact was ensured by a discrete constraint on the displacements in the vertical direction. This constraint was accomplished by a contact force, constrained to act vertically and only upward. The seat contact constraint contribution to the weak form was described by \u222b T 0 c4(t)\u03b4Zhip(q) dt (17) The discrete constraints were introduced by c4(t) > 0, (18) Zhip(q) = HFS, (19) where HFS is the footstretcher height, Fig. 1. 2.4 Fluid dynamic loads The fluid dynamic loads consist of hydrodynamic forces on the oar and boat, as well as air resistance on the rower, boat and oars. The air resistance force has been shown to be relatively small (up to 10 % of total drag for the system, [22]) compared to the hydrodynamic forces. The potential for reducing the air resistance by altering movement strategy was also expected to be very limited and, therefore, the air resistance forces were neglected in the present analysis, even if it is noted that they are strongly velocity-dependent"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003596_analsci.29.95-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003596_analsci.29.95-Figure1-1.png",
        "caption": "Fig. 1 Schematic illustration of the grafting of the GC electrode by the electrochemical reduction of in situ generated PV diazonium cation. (a) Five eqiuv. NaNO2, 1 M HCl, 4\u00b0C, (b) potential cycling between +0.6 and \u20130.8 V vs. Ag/AgCl. (c) Complexation of the metal ion with CA moiety on the electrode surface and the change in apparent charge of the CA complex.",
        "texts": [
            " Electrode grafting and further electrochemical measurements were conducted using an electrochemical analyzer (Model 650A, BAS, Japan) with a standard three-electrode configuration. A UV/Vis spectrophotometer (V-630, JASCO) was used for the measurements of absorption spectra of CA derivatives. A pH meter (Model F-52, Horiba) equipped with a combination glass electrode (9611-10D, Horiba) was calibrated by standard buffer solutions (pH 6.86, 4.01, and 1.68 at 25\u00b0C). Electrode grafting was performed by electrochemical reduction of an in situ generated diazonium cation13 as follows (Fig. 1). Pontacyl Violet 4BSN (PV) as a precursor, a CA derivative having an aromatic amine moiety with an azo linkage, was dissolved in a 1 M HCl solution at 4\u00b0C, then an ice-cold aqueous NaNO2 (5 equiv.) solution was added to form a diazonium cation of PV. After stirring of the mixture for 10 min at 4\u00b0C, a GC electrode was immersed into the mixture and the potential cycling (+0.6 to \u20130.8 V vs. Ag/AgCl (sat. KCl)) was carried out from the positive potential limit with a scan rate of 0.1 V s\u20131. Several grafting experiments under different conditions were performed at a series of potential cycle numbers (n) ranging from 1 to 30"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001736_978-3-319-01851-5_12-Figure12.6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001736_978-3-319-01851-5_12-Figure12.6-1.png",
        "caption": "Fig. 12.6 A two-dof rolling robot: (a) its general layout; and (b) a detail of its actuated wheels",
        "texts": [
            " Therefore, relations between these dependent and independent variables will be needed and will be derived in the course of our discussion. Moreover, we will study robots with both conventional and omnidirectional wheels. Of the latter, we will focus on robots with Mekanum wheels. We study here the robot of Fig. 10.17, under the assumption that it is driven by motors collocated at the axes of its two coaxial wheels, indicated as M1 and M2 in Fig. 10.17b. For quick reference, we repeat this figure here as Fig. 12.6. Our approach will be one of multibody dynamics; for this reason, we distinguish five rigid bodies composing the robotic mechanical system at hand. These are the three wheels (two actuated and one caster wheels), the bracket carrying the caster wheel, and the platform. We label these bodies with numbers from 1 to 5, in the foregoing order, while noticing that bodies 4 and 5, the bracket and the platform, undergo planar motion, and hence, deserve special treatment. The 6 6 mass matrices of the first three bodies are labeled M1 to M3, with a similar labeling for their corresponding six-dimensional twists, the counterpart items for bodies 4 and 5 being denoted by M0 4, M0 5, t0 4, and t0 5, the primes indicating 3 3\u2014as opposed to 6 6 in the general case\u2014mass matrices and three-dimensional\u2014as opposed to six-dimensional in the general case\u2014twist arrays",
            " First, we note that, from Eqs. (10.45), (10.52a and b), we can write, with ij denoting the .i; j / entry of \u201a, as derived in Sect. 10.5.1, !3 D . 11 P 1 C 12 P 2/e3 C \u0152 \u0131. P 1 P 2/C 21 P 1 C 22 P 2 k (12.43) or !3 D \u201a3 P a (12.44) with \u201a3 defined as \u201a3 D 11e3 C . 21 C \u0131/k 12e3 C . 22 \u0131/k In more compact form, \u201a3 D 11e3 C 21k 12e3 C 22k (12.45a) with 21 and 22 defined, in turn, as 21 21 C \u0131; 22 22 \u0131 (12.45b) Moreover, Pc3 D r P 3f3 D r. 11 P 1 C 12 P 2/f3 and hence, G3 D r 11f3 12f3 (12.46) Further, it is apparent from Fig. 12.6 that the scalar angular velocity of the bracket, !4, is given by !4 D ! C P and hence, !4 D \u0131. P 1 P 2/C 21 P 1 C 22 P 2 D 21 P 1 C 22 P 2 Therefore, we can write !4 D T4 P a (12.47a) where 4 is defined as 4 21 22 T (12.47b) Now, since we are given the inertial properties of the bracket in bracket coordinates, it makes sense to express Pc4 in those coordinates, taking into account that point C4 lies in the middle of the line PO3. Such an expression is obtained below: Pc4 D Po3 C !4 1 2 \u0152 d f3 C "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.14-1.png",
        "caption": "Fig. 2.14 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRPR (a) and 4RRPRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R||R||P||R (a) and R\\R||P||R||R||R (b)",
        "texts": [
            " 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2.1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003456_s0021894411020143-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003456_s0021894411020143-Figure1-1.png",
        "caption": "Fig. 1. Segment of the middle surface of the shell under internal pressure p.",
        "texts": [
            " \u2202 \u2202r (\u03bb2r) = \u03bb1 cos \u03b3, (4) where \u03bb1 = |dRs/dr| is the relative elongation of the middle surface in the meridional direction, \u03bb2 = \u03c1/r is the relative elongation of the middle surface in the circumferential direction, ur is the displacement of the middle plane in the initial state in the radial direction, uz is the displacement of the middle plane in the initial state along the Oz axis, and \u03b3 is the rotation angle of the normal to the middle surface. The body is deformed under a load of constant intensity p distributed over the inner surface and directed along the normal to the inner surface of the membrane (Fig. 1). The stress and strain distributions across the thickness of the shell is assumed to be homogeneous, i.e., the problem is considered in a membrane approximation. Using the Jourdain variation principle, we write the equilibrium condition for the membrane. Taking into account the incompressibility conditions, we have\u222b S p \u00b7 \u03b4V dS = \u222b V (\u03c3\u0303 \u00b7 \u00b7 \u03b4W\u0303 ) dV, where V is the volume of space occupied by the shell at an arbitrary time, S is the surface area acted upon by the forces p, \u03c3\u0303 is the stress tensor deviator, \u03b4V is the variation of the velocity field, and W\u0303 is the deviator of the strain rate tensor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003081_gt2012-69967-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003081_gt2012-69967-Figure1-1.png",
        "caption": "Figure 1. Representative motor shaft with six radial webs.",
        "texts": [
            " Torsional analyses can be conducted with coarse models that may not properly account for the mass-elastic properties of the train for all relevant modes, and therefore may not accurately predict torsional rotordynamic response. Modes associated with natural frequencies, particularly higher modes, can be missed with low fidelity models. Properly accounting for inertia and torsional stiffness properties is imperative to an accurate torsional analysis. This often includes creating a proper model of the complicated shaft geometry, such as webbed shafts, that are often seen in motors and generators. Synchronous and induction EMDs often have radial webs or spider bars that hold the windings or laminations. Figure 1 presents a generic motor shaft that includes six radial webs. Note that the laminations or windings that are attached to the outer portion of the webs are not shown in Fig. 1. Motor webs are often attached to the base shaft by shrink fits, keyways, or weldments. At times the shaft and webs are machined from the same forging. For welded or machined configurations, the radial webs contribute to the torsional stiffness of the shafting. The stiffening effect of keyed and shrink fit webs is generally considered less significant since the connections allow for some amount of slip. Additionally, motors can have other core configurations, such as slotted or wedged designs. While these motor designs also impact the torsional stiffness of the shafting, the bulk of the discussion for the current work will be focused on shaft designs that include fixed radial webs"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003721_gt2012-68476-Figure39-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003721_gt2012-68476-Figure39-1.png",
        "caption": "Figure 39 \u2013 Schematic Concept of U-Shaped Coupon Repair",
        "texts": [
            " The engine testing for an advanced coupon repair by diffusion brazing is successfully developed and the implementation is ongoing. This approach will also be described in next year's technical paper. U-Shaped Groove Coupon Diffusion Braze Repair As observed in fig 38, damage was evident on the U-shaped groove of the first cavity. The groove has a seal function, where the 1 st insert baffle fits into. Broken off pieces of this U-shaped groove was observed as this material for some reason was found to be very thin. The concept to repair this damage can be seen in figure 39. Figure 40 shows the wire EDM dovetail design for fixation of the coupon. Therefore there is a mechanical locking aspect combined with a diffusion braze repair to permanently hold in the coupon. Figure 41 shows the U-shaped coupon partially inserted in the dovetail slot and figure 42 shows the coupon fully inserted and braze paste applied to the periphery of the joint. A section of the U-shaped groove coupon repair is evident as seen in figure 43. 9 Copyright \u00a9 2012 by Alstom Technology Ltd. Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.14-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.14-1.png",
        "caption": "Figure 12.14 An airplane with six-DOF motions.",
        "texts": [
            "271) The first and third Euler equations are the equations of motion about the x- and z-axes: [ Mx Mz ] = [ I1\u03c9\u0307x I3\u03c9\u0307z ] (12.272) Therefore, a rolling rigid vehicle has a motion with four DOF which are translations in the x- and y-directions and rotations about the x- and z-axes. The Newton\u2013Euler equations of motion for such a rolling rigid vehicle in the body coordinate frame B are Fx = m v\u0307x \u2212 mr vy (12.273) Fy = m v\u0307y + mr vx (12.274) Mz = Iz\u03c9\u0307z = Izr\u0307 (12.275) Mx = Ix\u03c9\u0307x = Ixp\u0307 (12.276) Example 737 Motion of a Six-DOF Vehicle Consider a spacecraft or an airplane such as shown in Figure 12.14 that moves in space. Such a vehicle has six DOF. To develop the equations of motion, we define the kinematic characteristics as BvC = \u23a1 \u23a3 vx vy vz \u23a4 \u23a6 B v\u0307C = \u23a1 \u23a3 v\u0307x v\u0307y v\u0307z \u23a4 \u23a6 (12.277) B G\u03c9B = \u23a1 \u23a3 \u03c9x \u03c9y \u03c9z \u23a4 \u23a6 B G\u03c9\u0307B = \u23a1 \u23a3 \u03c9\u0307x \u03c9\u0307y \u03c9\u0307z \u23a4 \u23a6 (12.278) The acceleration vector of the vehicle in the body coordinate frame is Ba = B v\u0307B + B G\u03c9B \u00d7 BvB = \u23a1 \u23a3 v\u0307x + \u03c9yvz \u2212 \u03c9zvy v\u0307y + \u03c9zvx \u2212 \u03c9xvz v\u0307z + \u03c9xvy \u2212 \u03c9yvx \u23a4 \u23a6 (12.279) and therefore, Newtons equations of motion for the vehicle are \u23a1 \u23a3 Fx Fy Fz \u23a4 \u23a6 = m \u23a1 \u23a3 v\u0307x + \u03c9yvz \u2212 \u03c9zvy v\u0307y + \u03c9zvx \u2212 \u03c9xvz v\u0307z + \u03c9xvy \u2212 \u03c9yvx \u23a4 \u23a6 (12"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.64-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.64-1.png",
        "caption": "Fig. 2.64 4PaPaPR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\Pa\\\\P||R",
        "texts": [
            "22p) Idem No. 5 44. 4PPaRPa (Fig. 2.56) P||Pa\\R\\Pa (Fig. 2.22q) Idem No. 33 45. 4PPaPaR (Fig. 2.57) P||Pa||Pa\\R (Fig. 2.22r) Idem No. 4 46. 4PaPPaR (Fig. 2.58) Pa||P||Pa\\R (Fig. 2.22s) Idem No. 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002792_12.876079-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002792_12.876079-Figure4-1.png",
        "caption": "Figure 4. Principle of the pyrometer registering of the manufacturing process: t0 and tn are the start and end instants of the powder layer processing.",
        "texts": [
            " That is why the results of pyrometric measurements for wavelength \u03bb =1.26 \u00b5m are presented in arbitrary units in this paper. Proc. of SPIE Vol. 7921 79210D-2 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/29/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx For further discussion, it is necessary to introduce definition of a \u201ctrack\u201d and a \u201chatch distance\u201d. A track is the result of the laser beam scanning along a straight line on the powder bed with a constant speed. Hatch distance, \u03b4, (fig. 4d) is the distance between the neighbour tracks. To study SLM thermal phenomena, 10x10 mm2 surface area was subjected to laser beam scanning. The scanning strategy is presented in fig.4c: the scanning direction is the same within each layer. The tracks must be co-directional to avoid heat accumulation at their ends fig.4c. The laser beam jumps between the consecutive tracks with a much higher speed (vt12 = 7\u221910 3 mm/s) than the beam scanning speed (vt01 =120 mm/s) during track formation. Process parameters The experiments were carried out on Phenix PM100 machine (see SLM schematic in fig.1). Laser source is YLR-50 continuous wave Ytterbium fiber laser by IPG Photonics operating at 1075 nm wavelength with P = 50 W maximum output power. In the present study, 32 W power was delivered to the powder layer. The laser spot size on the surface of the powder bed is 70 \u00b5m in diameter"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure37-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure37-1.png",
        "caption": "Fig. 37 Chamfering Toe and Heel\u2014small pitch cone angle spiral bevel gear",
        "texts": [
            " 36, left, shows the Chamfer Tool with a 65\u02da Pivot Angle; there is still a positive angle between the cutting edge and the bottom of the tooth gap; by opposition, Fig. 36 right, the Pivot Angle is rather 45\u02da and clearly the movement must be stopped before the tool reaches the bottom of the tooth gap to avoid damaging the tool and creating a dent in the root. If the work piece has a small pitch cone angle, and this applies to cylindrical gears as well, then it is clear that while the Toe end of the tooth poses no issue, left Fig. 37, the Heel end cannot be chamfered with this tool because of collision risks between the tool spindle and the turn table, and also likely excessive turn table tilt\u2014right Fig. 37, unless the work holding support is long enough to allow sufficient clearance; otherwise, one must resort to using either End Mill, a Ball Mill or a CoSIMT tool. Of course, chamfering the tooth Tip causes no issue, either on the pinion or gear, as is shown in Figs. 38 and 39. the latter case, the cutting edges may be parallel to, or make an angle with the radial direction. Such tools come in several base geometries, and tool manufacturers such as Sandvik, P. Horn, Iscar, Ingersoll Rand make disk bodies to different diameters on which blade inserts are screwed"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003766_j.proeng.2011.04.543-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003766_j.proeng.2011.04.543-Figure1-1.png",
        "caption": "Fig. 1. (a) 1 dimensional model ; (b) 2 dimensional model ; (c) 3 dimensional model",
        "texts": [
            " / Procedia Engineering 10 (2011) 3291\u20133296 2 Author name / Procedia Engineering 00 (2011) 000\u2013000 view that two objects under impact carry out the same behavior as the case when contact surface is free. Therefore, in such conditions, if the natural frequency of a ball and a club head is the same, the restitution performance reaches the maximum. In order to find the indicator to the optimization of the golf clubhead, in this research, the impedance matching in the Multi-dimensional problems is verified numerically, and the boundary condition is also discussed. As shown in Fig. 1, the examinations of relation between restitution characteristic and eigen frequencies are performed with 1 dimensional model, 2 dimensional model and 3 dimensions model, respectively. Consider a body with an initial velocity of 40 m/s collides to another stationary body. For the sake of convenience, the impacting body is called \u201cclub\u201d, and the impacted body is called \u201cball\u201d blow. Numerical analyses are performed by LS-DYNA which is a kind of general purpose finite element software. Young\u2019s modulus of the club is changed from 1 to 30GPa",
            " The sensitivity function of rth eigen values can be derived as [4] (5) N i i N i N i i N i ii i m A m mv mv 1 1 10 1 22 1 42 ))(( , ( iiji r lkijkl r uuC )(()( , )) rrrr uuG Where is the rth eigen vector, and is the eigen value. )(ru )(r The procedure to create basis vectors is: 1) Undertake a modal analysis of design domain and select some of the eigenmodes. 2) Calculate the sensitivity functions of eigen values as the perturbation vector. 3) Exchange the perturbation vectors to the basis vectors of thickness variation. The analysis model is the same with 3 dimensional model in chapter 2 (see Fig.1 (a)). The Young's modulus of the club is set to 120GB which is the peak point of the coefficient of restitution. The design variable is the thickness distribution of the plate. About 15 eigenmodes are selected to generate the basis vectors, and some of them are shown in Fig. 6. Fig. 7 shows the optimized thickness distribution of the plate after 33 iterations. The thickness distribution of the plate is expressed by the color. It is found that the center area becomes thin and the outer area becomes thick"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001892_9781118354162.ch19-Figure19.8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001892_9781118354162.ch19-Figure19.8-1.png",
        "caption": "Figure 19.8 A catheter-type optical glucose sensor based on fluorescence quenching by oxygen. (1) Needle-type hollow container; (2) holes; (3) immobilized enzyme membrane; (4) optical fiber; (5) optical oxygen probe; (6) epoxy filling. Reprinted with permission from [27]. Copyright 2006 Elsevier.",
        "texts": [
            " As is typical of pH-linked enzymatic sensors, proper functioning requires the presence of a low buffer capacity pH buffer. As sensor functioning is based on pH variation, care should be exercised to avoid drastic effects of pH change on enzyme activity which can affect unfavorably the response function. Despite such problems, various applications of pH-linked optical transduction have been reported. For example, pH-based monitoring of acetylcholine esterase activity has proved useful in the development of inhibition-based sensors for pesticides and warfare agents. Figure 19.8 shows a glucose optical sensor based on fluorescence quenching by residual oxygen in the glucose oxidase-catalyzed reaction. This sensor is designed in the catheter format and has been used for either in vitro or in vivo determination of glucose in fish blood. Glucose oxidase was immobilized within a polymer film over an ultrathin dialysis membrane (3). The rolled enzyme membrane was place in a metal needle (1) with lateral holes that allow the sample to get in contact with the enzyme membrane. For transduction purposes, a fiber optic oxygen probe (4) is installed inside the membrane roll",
            " Depending on the mechanism of the enzymatic reaction, either hydrogen ions, oxygen, hydrogen peroxide, gases (ammonia) or the ammonium ion can serve for transduction by means of optical probes for such species. Recent advances in enzymatic optical sensors are surveyed in refs. [26,33]. The field of optical glucose sensors is reviewed in ref. [34]. Questions and Exercises (Section 19.4) 1 Review briefly the classes of enzyme that are suitable for producing optical enzyme sensors. For each class, indicate the reactant or product that can be detected by optical methods. 2 What are the main components of an optic fiber enzymatic sensor? 3 Comment on the role of each component of the sensor shown in Figure 19.8. 4 Draw a sketch of an electrochemiluminescence-based optical enzyme sensor. See ref. [28] for orientation. 5 Sketch possible configurations of optical urea sensors. 6 Discuss the configuration and the functioning mode of a chemiluminescence-based sensor using a dehydrogenase enzyme as the sensing element. 19.5 Optical Affinity Sensors Optical immunosensors can be designed in either the competitive or sandwich format. Owing to its outstanding sensitivity, fluorescence is the preferred optical transduction method"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.19-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.19-1.png",
        "caption": "Fig. 2.19 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRC (a) and 4RCRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R||R||C (a) and R\\C||R||R (b)",
        "texts": [
            " 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No. 19 28. 4PRRRP (Fig. 2.15b) P\\R||R\\R\\P (Fig. 2.1z1) Idem No. 19 29. 4PRRPR (Fig. 2.16) P\\R\\R\\P\\kR (Fig. 2.1a0) Idem No. 5 30. 4CRRR (Fig. 2.17a) C\\R||R\\R (Fig. 2.1d0) Idem No. 1 (continued) Table 2.1 (continued) No. PM type Limb topology Connecting conditions 31. 4CRRR (Fig. 2.17b) C||R||R\\R (Fig. 2.1e0) Idem No. 13 32. 4RRCR (Fig. 2.18a) R||R||C\\R (Fig. 2.1f0) Idem No. 13 33. 4RCRR (Fig. 2.18b) R||C||R\\R (Fig. 2.1g0) Idem No. 13 34. 4RRRC (Fig. 2.19a) R\\R||R||C (Fig. 2.1h0) Idem No. 21 35. 4RCRR (Fig. 2.19b) R\\C||R||R (Fig. 2.1l0) Idem No. 21 36. 4RRCR (Fig. 2.20a) R\\R||C||R (Fig. 2.1b0) Idem No. 21 37. 4RCRR (Fig. 2.20b) R||C\\R||R (Fig. 2.1c0) Idem No. 21 j\u00bc1 fj 5 5 23. Pp2 j\u00bc1 fj 5 5 (continued) In the fully-parallel topologies of PMs with coupled Sch\u00f6nflies motions F / G1G2-G3-G4 presented in this section, the moving platform n : nGi (i = 1, 2, 3, 4) is connected to the reference platform 1 : 1Gi : 0 by four spatial complex limbs with four or five degrees of connectivity. The complex limbs combine only revolute, prismatic and cylindrical joints"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001736_978-3-319-01851-5_12-Figure12.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001736_978-3-319-01851-5_12-Figure12.5-1.png",
        "caption": "Fig. 12.5 The serial manipulator of the J th leg",
        "texts": [
            " In this figure, the legs have been replaced by the constraint wrenches fwC J gVII acting at point CM, the governing equation thus taking the form of Eq. (7.5c), namely, MMPtM D WMMMtM C wW C VIX JDI wC J (12.11) with wW denoting the external wrench acting on M. Furthermore, let us denote by qJ the variable of the actuated joint of the J th leg, all variables of the six actuated joints being grouped in the six-dimensional array q, i.e., q qI qII qVI T (12.12) Now, we derive a relation between the twist tM and the active joint rates, PqJ , for J D I , II , : : :, VI . To this end, we resort to Fig. 12.5, depicting the J th leg as a serial-type, six-axis manipulator, whose twist\u2013shape relations are readily expressed as in Eq. (5.9), namely, JJ P J D tM; J D I; II; : : : ; VI (12.13) where JJ is the 6 6 Jacobian matrix of the J th leg. In Fig. 12.5, the moving platform M has been replaced by the constraint wrench transmitted by the moving platform onto the end link of the J th leg, wC J , whose sign is the opposite of that transmitted by this leg onto M by virtue of Newton\u2019s third law. The dynamics model of the manipulator of Fig. 12.5 then takes the form IJ R J C CJ . J ; P J / P J D J JTJ wC J ; J D I; II; : : : ; VI (12.14) where IJ is the 6 6 inertia matrix of the manipulator, while CJ is the matrix coefficient of the inertia terms that are quadratic in the joint rates. Moreover, J and J denote the six-dimensional vectors of joint variables and joint torques, namely, J 2 6664 J1 J2 ::: J6 3 7775 ; J 2 6666666664 0 ::: Jk 0 ::: 0 3 7777777775 (12.15) with subscript Jk denoting in turn the only actuated joint of the J th leg, namely, the kth joint of the leg",
            "16) If the actuated joint is prismatic, as is the case in flight simulators, then fJ is a force; if this joint is a revolute, then fJ is a torque. Now, since the dimension of q coincides with the degree of freedom of the manipulator, it is possible to find, within the framework of the natural orthogonal complement, a 6 6 matrix LJ mapping the vector of actuated joint rates Pq into the vector of J th-leg joint-rates, namely, P J D LJ Pq; J D I; II; : : : ; VI (12.17) The calculation of LJ will be illustrated with an example. Moreover, if the manipulator of Fig. 12.5 is not at a singular configuration, then we can solve for wC J from Eq. (12.14), i.e., wC J D J T J . J IJ R J CJ P J / (12.18) in which the superscript T stands for the transpose of the inverse, or equivalently, the inverse of the transpose, while IJ D IJ . J / and CJ D CJ . J ; P J /. Further, we substitute wC J as given by Eq. (12.18) into Eq. (12.11), thereby obtaining the Newton\u2013Euler equations of the moving platform free of constraint wrenches. Additionally, the equations thus resulting now contain inertia terms and joint torques pertaining to the J th leg, namely, MMPtM D WMMMtM C wW C VIX JDI J T J "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure29-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure29-1.png",
        "caption": "Fig. 29 End Mill installation for Toe/Heel chamfering\u2014helical and spiral bevel gears",
        "texts": [
            " 27); and if the Pivot Angle is correctly chosen, tool interference with the tooth flank can be avoided (right, Fig. 27). The Ball Mill tool (BM), thanks to its spherical end, can be fitted in places where an End Mill tool would not do an acceptable job. Consider for example the spiral bevel pinion shown in Fig. 28, left. The Ball Mill tool can be plunged vertically along the Toe and Heel edges without any risk of tool spindle to turn table interference. And by carefully selecting the Ball Mill diameter, the fillet area can also be chamfered. Five unit vectors (left, Fig. 29) are required to control the BM at any point along a tooth edge: Vo: N X T ,\u2212\u2212\u2192 Tool: the tool vector,\u2212\u2212\u2212\u2192 Trans: axis about which vector \u2212\u2212\u2192 Tool is rotated by the Pivot Angle. \u2212\u2192 Vo is obtained from the cross product of N and T . Again, vector \u2212\u2212\u2212\u2192 Trans is obtained by pivoting T about \u2212\u2192 Vo by \u03c0 2 + the local pressure angle; vector \u2212\u2212\u2192 Tool is obtained from the cross product of \u2212\u2212\u2212\u2192 Trans and \u2212\u2192 Vo; vector \u2212\u2212\u2192 Tool can be pivoted parallel to itself about the local tooth edge point by the Chamfer Angle\u2014left, Fig. 29\u2014and about vector \u2212\u2212\u2212\u2192 Trans by the user inputted Pivot Angle. However, there is a potential issue with this type of tool in that if the diameter of the spherical end is small, which is required to fit in the fillet area, a double lip is created where the tool enters and exits the tooth edge, as Fig. 30 shows. In many instances, this is unacceptable and either an alternative solution is required, or else the BM diameter is increased to minimize the double lip and the fillet area is not targeted. Chamfer Tools (CT) are widely available and come in varied sizes"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003488_kem.518.76-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003488_kem.518.76-Figure1-1.png",
        "caption": "Fig. 1. Gear model.",
        "texts": [
            " Relatively recently, in the quest for accurate time and frequency resolution, Huang et al. [27] proposed the Empirical Mode Decomposition method (EMD). Since then, attention was gained in applying EMD in the damage detection of gears [28, 29]. This technique decomposes the signal into intrinsic mode functions and the instantaneous frequency and amplitude of each intrinsic mode function can be then obtained by applying the Hilbert Transform. Theoretical model. In the present work a simplified two-degree-of-freedom (2DOF) torsional model is used, as shown in Fig. 1. The gearbox shafts and bearings are assumed to be rigid. The model was chosen because of its strongly nonlinear characteristics. This nonlinear model has been tested in previous research work [8, 10, 12]. The dynamics of the system can be described by a 2DOF system, with coordinates: \u0305 = , (1) The equations of motion describing the gear model given in the general form are: + , , = , , (2) \u2212 , , = \u2212 , , (3) This model takes into account the static transmission error representing the geometrical errors of the gear teeth profile and spacing, "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002989_2425296.2425316-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002989_2425296.2425316-Figure3-1.png",
        "caption": "Figure 3: Spherical joint at the shoulder.",
        "texts": [
            " And the potential energy component, \u03b2 \u2211 mig T r\u03040i , considers the en- ergy needed for maintaining the desired pose. By adjusting the weighting factors \u03b1 and \u03b2, the minimization of the objective function will yield different results. For example, a large \u03b2 value suggests that in the final posture, the elbow should be kept low. A large \u03b1 value, however, usually means that the joint movements will be as little as possible. As discussed in the earlier section, the position of the elbow is defined by joints q1 and q2. As illustrated in Figure 3, let (x,y,z) be the the elbow position with respect to a local coordinate frame (Ox, Oy , Oz ) located at the shoulder joint. By geometry, it appears that the joint angles q1 and q2 can be computed as follows: q1 = tan\u22121 ( y x ) q2 = tan\u22121 (xr z ) (4) where xr = x cos q1 + y sin q1 (5) is the x-coordinate of the new frame after the original frame is rotated by q1 about the z-axis. However, due to the fact that atan2(\u00b7) \u2208 (\u2212180\u25e6, 180\u25e6], there may be discontinuous jump in the solution (4) when crossing \u00b1180\u25e6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003302_amr.487.327-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003302_amr.487.327-Figure12-1.png",
        "caption": "Fig. 12 The filling process: STEP=800",
        "texts": [
            " After that, the materials, contact conditions, boundary conditions, gravity, initial conditions and the operating parameters are defined respectively to complete the pre-treatment of the numerical simulation of the filling and the solidification process of the air intake hood [4][5][6]. Then, all the defined parameters were saved and the dynamic simulation of the filling and solidification process of the air intake hood could be operated. Figure 7, Figure 8, Figure 9, Figure 10, Figure 11 and Figure 12 are the results of the FEM simulation of the filling and solidification conditions at different moments. It can be seen through the simulation of the casting filling and solidification process of the air intake hood that the alloy liquid was poured into the bigger hole of the air intake hood first. Then the rest began to be filled after the big hole was full (STEP=100), until all parts were full. Because of the contract effect with the plaster model, the temperature of the alloy liquid in the front of the air intake hood (the location of the thin-walled away from the gate) was decreased rapidly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.40-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.40-1.png",
        "caption": "Fig. 2.40 4PPaPaR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology P\\Pa\\\\Pa\\||R",
        "texts": [
            "21t) Idem No. 5 21. 4PaPRP (Fig. 2.36a) Pa\\P||R\\P (Fig. 2.21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.42-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.42-1.png",
        "caption": "Fig. 3.42",
        "texts": [
            "41c) M = \u221a M2 2 + M2 3 = \u03981 \u03c92 0 R r . Its direction is horizontal and perpendicular to the axis of the wheel. The compression force N between the wheel and the crushing bed is given by the sum of the weight W of the wheel and the gyroscopic part M/R (Fig. 3.41d): N = W + M R = W + \u03981 \u03c92 0 r . Therefore, the compression force can be considerably increased by increasing \u03c90. E3.19Example 3.19 Determine the moments and products of inertia with respect to the axes x\u0304, y\u0304, z\u0304 and x, y, z for the homogeneous cuboid of mass m shown in Fig. 3.42. x\u0304 Solution With dm = \u03c1 dx\u0304dy\u0304 dz\u0304 and m = \u03c1 abc the quantities \u0398x\u0304 and \u0398x\u0304y\u0304 are obtained as \u0398x\u0304 = \u222b (y\u03042 + z\u03042) dm = \u03c1 c\u222b 0 b\u222b 0 a\u222b 0 (y\u03042 + z\u03042) dx\u0304 dy\u0304 dz\u0304 = m 3 (b2 + c2), \u0398x\u0304y\u0304 = \u2212 \u222b x\u0304y\u0304 dm = \u2212 \u03c1 c\u222b 0 b\u222b 0 a\u222b 0 x\u0304y\u0304 dx\u0304dy\u0304 dz\u0304 = \u2212 mab 4 . Analogously we find \u0398y\u0304 = m 3 (c2 + a2), \u0398z\u0304 = m 3 (a2 + b2), \u0398y\u0304z\u0304 = \u2212mbc 4 , \u0398z\u0304x\u0304 = \u2212mca 4 . Because x, y, z are axes of symmetry, they are principal axes. Therefore, the products of inertia \u0398xy, \u0398yz, \u0398zx are zero. The principal moments of inertia are given by \u03981 = \u0398x = 8 \u03c1 c/2\u222b 0 b/2\u222b 0 a/2\u222b 0 (y2 + z2) dxdy dz = m 12 (b2 + c2), \u03982 = \u0398y = m 12 (c2 + a2), \u03983 = \u0398z = m 12 (a2 + b2) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003431_2013-36-0272-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003431_2013-36-0272-Figure7-1.png",
        "caption": "Figure 7 \u2013 Torque applied in the wheel",
        "texts": [
            " Expressing in radians, the \u201cstrain relief\u201d (negative) will be: (22) with: and (23) where a is the vertical distance of wheel center O to the ground, in the deformed tire. Using equations (22) in equation (12), we obtain: (24) where is given by equation (23). Then, the relationship between the wheel\u2019s angular velocity and its center linear velocity Ve depends on the factor sin / , which is a function of the vertical tire deflection (R \u2013 a), and this deflection depends of the applied vertical load and the tire internal pressure p. This case corresponds to the constant acceleration case seen before. Let us consider the situation shown in Figure 7, where the point O is moving to the right and the torque T is applied as shown. The force acting in the plane horizontal contact region of the tire belt, at section B shown in Figure 7 (b), is (F \u2013 F).cos . It means that there is an additional \u201cstress relief\u201d and corresponding \u201cstrain relief\u201d in this section of the tire belt region, in contact with the ground. When the wheel rotates, as there is no sliding between the ground and the tire surface, this deformation will remain the same through the entire extension of the contact region. Page 6 of 7 From Figure 7 (a), using point O as the pole and neglecting the rotational inertias, we have for the moments: (25) Using the plane stress theory [5] for the cylindrical surface of the tire belt, supposing that the torque T will not affect the axial (z-direction) stress component, and using an equivalent Young\u2019s modulus E (neglecting the fact that the tire belt material is an anisotropic composite), for the circumferential direction we have: (26) with Now, using equations (18), (25) and (1), we obtain the relationship between the angular velocity of the wheel and the linear velocity V2 of its center O: (27) where: R = external radius of the unloaded tire; = arc cos (a / R), where a is the vertical distance of wheel center O to the ground, in the deformed tire; T = torque applied to the wheel; E = equivalent Young\u2019s modulus of tire material, in the circumferential direction (noting that this material is an anisotropic composite); A = equivalent cross section area of the tire belt (which depends on tread grooves design)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002655_amr.789.443-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002655_amr.789.443-Figure4-1.png",
        "caption": "Figure 4: LCS with Battery Figure 5: LCS with Battery and Lamps",
        "texts": [
            " To get the relation between the efficiency of electric generator versus rotation per minute (rpm) on 24 volt voltage, same experiments were performed without load on voltage about 25 volt up to about 31 volt. The load was controlled in order to prevent the electric generator produced voltage less than 20 volt. To get the relation between battery charging current and speed, the data needed were battery voltage, electric current, electric generator voltage, as well as electric generator speed and torque. Fig. 4 showed the scheme. The experiment on baterry charging was performed for low, normal, and full condition as well. Initial voltage of low battery condition was 8.1 V, while the initial voltage of normal battery condition and full battery condition were 12,10 V and 13,44 V respectively. The electric generator speed was controlled until produced electric voltage about 10 V and then was connected to the battery. All data were observed and recorded. The battery was then disconnected from the system. Same experiments were done for electric voltage about 11 V up to about 26 volt"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001767_9781118316887.ch8-Figure8.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001767_9781118316887.ch8-Figure8.2-1.png",
        "caption": "Figure 8.2-3: Two-pole, three-phase, 28-V, 0.63-hp, 4500-r/min, permanent-magnet ac machine (courtesy of Vickers Electromech).",
        "texts": [
            " We will find that after all these new concepts are out of the way, it is then rather easy for us to step up to the three-phase device. If you have studied Chapter 7 on synchronous machines, you will be able to move rapidly through the introductory material. However, do not feel that you need to study the material in Chapter 7 to follow the presentation in this chapter. It is written assuming that the reader does not have a background in synchronous machines. A two-pole two-phase permanent-magnet ac machine is shown in Fig. 8.2-1. The stator windings are identical, sinusoidally distributed windings each with Ns equivalent turns and resistance rs as described in Chapter 4. The angular 8.2. TWO-PHASE PERMANENT-MAGNET AC MACHINE 347 displacement about the stator is denoted 0S, referenced to the as axis. The angular displacement about the rotor is 0r, referenced to the q axis. The angular velocity of the rotor is uor and 6r is the angular displacement of the rotor measured from the as axis to the q axis. Thus, a given point on the rotor surface at the angular position 0r may be related to an adjacent point on the inside stator surface with angular position (j)s as (j)s = (j)r + er (8.2-1) The d axis is fixed at the center of the north pole of the permanent-magnet rotor and the q axis is displaced \\ir counterclockwise from the d axis. The electromechanical torque Te and the load torque TL are also indicated in Fig. 8.2-1. As defined in Chapter 2, Te is assumed positive in the direction of increasing 6r; TL is positive in the opposite direction. In the following analysis, it is assumed that: 1. The magnetic system is linear. 2. It is assumed that the q- and d-axis reluctances are equal. Unequal reluctance is treated briefly in Section 8.11. The slight rotor indentation in the q axis is shown in Fig. 8.2-1 to indicate separation of the N and S poles. 3. The open-circuit stator voltages, induced by rotating the permanentmagnet rotor at a constant speed, are sinusoidal. 4. Large stator currents can be tolerated without significant demagnetization of the permanent magnet. 5. Damper windings (short-circuited rotor windings) are not considered. Although these assumptions may be somewhat oversimplifying, this type of analysis is convenient while still portraying the main operating features of the device",
            " Hence, those who have studied the material in Chapter 7 may find most of the material in the next two sections a review, offering little challenge. In this chapter, as in previous chapters, we will consider the two-phase machine before the three-phase counterpart for convenience of analysis. Even though most brushless dc machines are three-phase devices, from an analytical standpoint, it is to our advantage to consider the two-phase device and then extend our work to the three-phase system. In Fig. 8.2-1, the magnetic axes of the stator windings are denoted as the as and bs axes. The d axis (direct axis) is used to denote the magnetic axis of the permanent-magnet rotor and the q axis (quadrature axis) is used to denote an axis \\K ahead of the d axis. The concept of the q axis and d axis is reserved for association with the rotor magnetic axes of synchronous machines since, over the years, this association has become convention. Electromagnetic torque is produced by the interaction of the poles of the permanent-magnet rotor and the poles resulting from the rotating airgap mmf established by currents flowing in the stator windings. The rotating mmf (mmfs) established by symmetrical two-phase stator windings carrying balanced two-phase currents is given by (4.4-11). The two sensors shown in Fig. 8.2-1 may be Hall-effect devices. When the north pole is under a sensor, its output is nonzero; with a south pole under the sensor, its output is zero. The stator of the permanent-magnet ac machine is supplied from a dc-to-ac inverter, the frequency of which corresponds to the rotor speed. The states of the sensors are used to determine the switching logic for the inverter, which, in turn, determines the output frequency of the inverter. In the actual machine, the sensors are not positioned over the rotor as shown in Fig. 8.2-1. Instead, they are placed over a ring that is mounted on the shaft external to the stator and magnetized as the rotor. We will return to these sensors and the roles they play later. A four-pole, three-phase, 28-V, \u00a7-hp, permanent-magnet ac machine is shown in Fig. 8.2-2. The disassembled motor is shown in Fig. 8.2-2a, wherein the stator windings are visible. The opposite end of the stator housing is shown in Fig. 8.2-2ft. Housed therein are the Hall-effect sensors, which are used to determine rotor position, the drive inverter, the filter capacitor, and 350 PERMANENT-MAGNET ac MACHINE the logic circuitry. The stator and rotor of a ten-pole, three-phase, 28-V, 0.63-hp, 4500-r/min, permanent-magnet ac machine is shown in Fig. 8.2-3. The magnets are samarium cobalt and the drive inverter is supplied from a 28-V dc source. The magnetic end cap is used in conjunction with Hall-effect sensors mounted in the stator housing (not shown) to determine the rotor position. SP8.2-1 Express mmfr for the two-pole, two-phase, permanent-magnet ac machine shown in Fig. 8.2-1. Let Fp denote the peak value. [mmfr = -Fp sin(4>s - er)} WINDING INDUCTANCES OF A PERMANENT-MAGNETIC ac MACHINE The voltage equations for the two-pole, two-phase, permanent-magnet ac machine shown in Fig. 8.2-1 may be expressed as Vas = rsias + \u2014^- (8.3-1) Vbs = rsibs + \u2014^- (8.3-2) In matrix form, Vabs = T siabs + pKbs (8.3-3) where p is the operator d/dt, and for voltages, currents, and flux linkages, (fabs)T = [fas fbs] (8.3-4) with rs 0 0 r\u00ab (8.3-5) A review of matrix algebra is given in Appendix B. The flux-linkage equations may be expressed as Aas = -^asas^as i -^asbs^bs T\" \u0302 asm \\0.o-\\)) ^bs \u2014 Lbsasias + L^^sHs + Xbsm (8.3-7) In matrix form, \\bs = LS\u00eea6s + Am (8.3-8) 352 PERMANENT-MAGNET ac MACHINE where XL is the column vector: A, A\u2122 xasm ^bsm = K m sin 9r \u2014 COS 6r (8.3-9) In (8.3-9), \\'m is the amplitude of the flux linkages established by the permanent magnet as viewed from the stator phase windings. In other words, the magnitude of A^ is proportional to the magnitude of the open-circuit sinusoidal voltage induced in each stator phase winding. It may be helpful to visualize the permanent-magnet rotor as a rotor with a winding carrying a constant current and in such a position to cause the north and south poles to appear as shown in Fig. 8.2-1. The rotor displacement 6r is expressed as CiUr ~^=^r (8.3-10) We will assume that the reluctance of the rotor of the permanent-magnet ac machine is the same in the q- and d-axes. Actually, this assumption may be an oversimplification in some cases; however, it markedly reduces our work and allows us to establish directly the basic principles of the controlled permanent-magnet ac machine without significant error. Unequal q- and daxes reluctances is considered briefly in Section 8.11 and is taken into account in more detail in [1]",
            " The actual rotor speed at which the measurement was taken is _ (r/min) (rad/r) s/mm (1000)(2TT) 100 = V 6Q ^ = -g-*r rad/s (8A-1) Prom (4.5-10), the electrical angular velocity is _ P 4 IOOTT 200 J . / n A rt, = 2^~ = ~Tn r a d / s (8A\"2) With the phases open-circuited, ias = ibs \u2014 0. Thus, from (8.3-1) and (8.3-9), d(\\'m sin Br) dt = \\'murcos9r (8A-3) Now the peak-to-peak voltage is 34.6 V; hence, from (8A-3), with the peak-to-peak voltage divided by 2, we have 34.6 w ,200 , ,nA \u201e, - y = Kni-Y*) (8A-4) Solving for \\'m yields SP8.3-1 The stator windings of the permanent-magnet ac machine shown in Fig. 8.2-1 are open-circuited. The rotor is driven clockwise and Vas = -10 sin 100*. Determine V^. [Vbs = - 1 0 coslOOt] SP8.3-2 Determine A^ for SP8.3-1. [A^ = 0.1 V \u2022 s/rad] SP8.3-3 During steady-state operation, (8.3-10) becomes 6r = u>rt + 0r(O). What is 0r(O) in SP8.3-1? [0r(O) = \u2014 |TT] 354 PERMANENT-MAGNET ac MACHINE An expression for the electromagnetic torque may be obtained by using the second entry in Table 2.5-1. Since we are assuming a linear magnetic system, the coenergy may be expressed as Wc = -Lssf\u00e2s + i2 hs) + Xf mias sin9r - A^i6s cos6r + Wpm (8.4-1) where Wpm is the energy associated with the permanent magnet, which is constant for the device shown in Fig. 8.2-1. Taking the partial derivative with respect to 6r yields p Te = -W^m^as COS 6r + lhs SHI 6r) (8.4-2) The above expression is positive for motor action. The torque and speed may be related as Te = j ( | ) ^ + \u00df m ( | ) u , r + TL (8.4-3) where J is in kg \u2022 m2; it is the inertia of the rotor and the connected mechanical load. Since we will be concerned primarily with motor action, the load torque TL is assumed positive, as indicated in Fig. 8.2-1. The constant Bm is a damping coefficient associated with the rotational system of the machine and mechanical load. It has the units N \u2022 m \u2022 s/rad and it is generally small and often neglected in the case of the machine but may be considerable for the mechanical load. SP8.4-1 Calculate Te for a two-pole permanent-magnet ac motor if ias = cos 0r, and %s = sm6r. \\'m = 0.1 V \u2022 s/rad. [Te = 0.1N- m] SP8.4-2 Repeat SP8.4-1 with 9r = urt and ias = cos(\u00fc\u00fcrt + \\-K) and ibs \u2014 sin(a;rt+|7r). [Tc = 0] PERMANENT-MAGNETIC ac MACHINE IN THE ROTOR REFERENCE FRAME For the permanent-magnet ac machine, the reference frame fixed in the rotor is the reference frame of choice",
            " With the rotor speed and position available, the dc-to-ac inverter is controlled so that the frequency of the fundamental component of the voltages applied to the stator windings is equal instantaneously to the electrical angular velocity of the rotor ujr. Also, the switching of the inverter can be advanced or retarded, thereby shifting the phase or time position of the phase voltages by changing 9esv relative to the rotor position 9r. In the case of the brushless dc machine 9esv is shifted such that (frv (8.5-16) is zero, whereupon the peak positive values of vas occur at every instant the q axis (Fig. 8.2-1) is horizontal and directed to the right. With (j)v \u2014 0, vqs = \\/2vs and vjs \u2014 0. This angular relationship between the stator voltages and the magnetic poles of the rotor is tightly maintained by control of the inverter during steady-state operation as well as during transient changes in rotor speed ujr due to changes in torque load TL and/or the dc voltage of the inverter vs. Let us explain the action of the dc-to-ac inverter in a slightly different way. To do this, let us fix time at the instant vas is at its peak value. If we looked at the q axis at this instant, it would be horizontal and to the right, whereupon, the N pole would be at (j)r = \u2014 \\K and the S pole at (j)r = \\K (Fig. 8.2-1). Now with <pv = 0 and for steady-state operation, Vas \u2014 V%VS cos c\u00fcrt and 9r = cjri, and we have fixed time at zero in order to portray the situation described above, where the positive peak value of Vas 358 PERMANENT-MAGNET ac MACHINE occurs when the q axis is horizontal to the right. Now let time proceed and when 9r = 27r the q axis is again horizontal and to the right and Vas has also advanced 2TT and it is again at its positive peak. Ideally, the dc-to-ac inverter is designed to maintain this relative position between the rotor and the peak value of vas during transient and steady-state operation",
            "8-8) Fbs = V2FS sm[ujrt + 9esf(0)} (8.8-9) From (5.5-9) and (5.5-10) with u = ujr and 6(0) = #r(0), Fr qs = V2FS cos[9esf(0) - 9r(0)} (8.8-10) Fr ds - -V2FS sin[Oes/(0) - 0r(O)] (8.8-11) The phasor that represents Fas given by (8.8-8) is Fas = F8\u00e9>9\"fW (8.8-12) which may be written as Fas = Fs cos 6esf(0) + jFs sin 0^/(0) (8.8-13) Now, since we can choose our time zero as we wish, let us select it so that 0r(O) = 0. In other words, our time zero will always be selected at the time the q axis in Fig. 8.2-1 is horizontal to the right. For those of you who have read Chapter 7 on synchronous machines, you will note a difference; there we selected time zero so that 9e8V(0) \u2014 0. If 0r(O) = 0, then Fr qs, (8.8-10) and -FT ds, (8.8-11) are, respectively, the same as V2 times the real and imaginary parts of (8.8-13). Thus, for 0r(O) = 0, V2Fas = Fr qs-jFrds (8.8-14) 368 PERMANENT-MAGNET ac MACHINE As in Chapter 5 and Example 8B, we have equated a phasor, which represents a sinusoidal quantity, to Fqs and Fds, which are constants"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure10-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure10-1.png",
        "caption": "Figure 10: Prototype system of Hohlraum assembly",
        "texts": [
            " From Figure 8, we can see the geometric boundaries of those subassemblies which are traced real-time by the detection software. The TMP and Hohlraum manipulation stages will keep moving until the geometric boundaries of TMP and Hohlraum turn to be coincidence. Figure 8: Components of optical image system Figure 9: Geometric boundaries of the subassemblies can be traced real-time 4. RESULTS According to the design of precision robotic assembly system, the prototype system is developed. In the prototype system, the manual stages are used to substitute for the auto liner stages (Figure 10). There are three detections (shade outline detection, coaxial position detection and force / torque detection) in the prototype system (Figure 11). Source of parallel light Keyence laser sensor CCD Source of coaxial light Long-focus lenCatoptron mirror Auto liner stages Proc. of SPIE Vol. 8418 841819-4 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 04/21/2013 Terms of Use: http://spiedl.org/terms The first experiment which is carried out on Hohlraum inserted into TMP by the prototype system is successful, and the finished Half Hohlraum Component is shown in Figure 13"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002833_imece2013-62921-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002833_imece2013-62921-Figure2-1.png",
        "caption": "Figure 2: a) Multiple pined connection under bending. b) Internal forces in the connection. Shear forces on each pin have two components ( \u2032 , \" ) which are created by the shear force (F) and the moment (M) respectively.",
        "texts": [
            "org/about-asme/terms-of-use a b Figure 1: The mechanical optimization process using two methods. a) Coupling finite element software with an optimization tool. b) Using mathematical model. There are various types of bolted connections which can be studied as an optimization case for the layout of connectors. Here are two metal members, similar in material and thickness, are connected by several pins. The underneath member is fix supported in one side, and the upper one is under bending by a force which is normal to the fixed support direction and applied to the member end (Figure 2.a). The pins prevent the member from moving by being under shear forces. Figure 2.b shows the internal forces created in the connection areas which their results are applied to the center of pins` area (O). Shear forces applied to each pin are also shown in the Figure. The shear force applied to each pin ( ) is composed from two components: the resulting shear force created by uniform shear load (primary load ( \u2032 ) distribution on pins which is calculated for each one from equation (1), and the resulting force from the moment (M) (secondary shear force) which is proportional to the distance between the center of each pin and the central point of pins ( )"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002485_s1068366613060123-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002485_s1068366613060123-Figure1-1.png",
        "caption": "Fig. 1. Equilibrium of solid during rolling.",
        "texts": [
            "3103/S1068366613060123 JOURNAL OF FRICTION AND WEAR Vol. 34 No. 6 2013 STATIC COEFFICIENT OF ROLLING FRICTION UNDER HEATING 451 coefficient of rolling friction are available in the litera ture, information on the effect of heating is lacking. The aim of this work was to study the influence of heating on the static coefficient of rolling friction. The rolling of a ball over a flat surface is described by the following equations of static equilibrium: (1) (2) In order to determine the force of rolling friction (Fig. 1a), from (1) and (2) we obtain the following expressions: N \u2013 G = 0; F \u22c5 r \u2013 N \u22c5 e = 0; (3) hence, (4) where r is the radius of the ball, G is the weight of the ball, and e is the distance from point A, at which the resultant reaction N is applied. In the case when the ball rolls downward over the inclined plane (Fig. 1b), the drive moment is equilibrated by the moment of resistance Expression (4) also follows from this. The application of expression (2) to the boundary case of equilibrium on the inclined plane (Fig. 1b) yields the following relation: (5) or (6) where \u03b1 is the angle of inclination of the plane and \u03bcR is the static coefficient of friction. 0;xF =\u2211 0,yF =\u2211 0.AM =\u2211 ,e eF G N r r = \u22c5 = \u22c5 RM F r= \u22c5 .M N e\u00b5 = \u22c5 sin cos 0,G r G e\u03b1 \u22c5 \u2212 \u03b1 \u22c5 = tan ,R e r = \u03b1 = \u03bc Experiments were carried out using a tribometer with an inclined plane (Fig. 2) [1]. In this tribometer, contact pairs, along with a heater and a temperature sensor that are located very near to the zone of con tact, can rotate at a required angle \u03b1 with regard to the horizontal plane"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000764_1-4020-5123-9_3-Figure3-9-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000764_1-4020-5123-9_3-Figure3-9-1.png",
        "caption": "Figure 3-9. The 1DPath model and the adjustable geometry parameters.",
        "texts": [
            " Then, after starting the simulation, the control window (top right) opens, which allows interactive control of the simulation process. Each template set provides its own parameter and control panel. Currently, three models are implemented; extending the library is easily done by extending the Simulation Java class. Details of the available models and the Java class 67 Chapter 3 are provided in the user manuals [Lienemann et al., 2004a, Lienemann et al., 2004b]. The first model (1DPath) provides a line of electrode pads both for confined and non-confined droplets (Fig. 3-9). It allows one to test basic operations of an electrowetting array like moving, dispensing, merging and splitting. Since the topology of the droplet remains unchanged during the simulation, splitting and merging is detected by the designer using the graphical output. This model is also useful to explore the exact droplet shape, which is, e.g., valuable for estimation of optical properties for micro-lens design [Berge and Peseux, 2000], and how the shape behaves during the basic operations. This may also be critical for the optical detection of fluorescently marked biological molecules where refraction effects and signal variation due to the droplet depth must be compensated"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.30-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.30-1.png",
        "caption": "Fig. 2.30 4PPPaR-type fully-parallel PMs with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology P\\P\\kPa\\kR (a) and P\\P||Pa\\\\R (b)",
        "texts": [
            "21c) The second and the last joints of the four limbs have parallel axes 4. 4PPPaR (Fig. 2.27b) P\\P\\kPa\\R (Fig. 2.21d) The last revolute joints of the four limbs have parallel axes 5. 4PRPPa (Fig. 2.28a) P||R\\P||Pa (Fig. 2.21e) The second joints of the four limbs have parallel axes 6. 4PPRPa (Fig. 2.28b) P\\P||R\\Pa (Fig. 2.21f) The third joints of the four limbs have parallel axes 7. 4PRPPa (Fig. 2.29a) P\\R||P\\Pa (Fig. 2.21g) Idem No. 5 8. 4PRPaP (Fig. 2.29b) P\\R\\Pa \\kP (Fig. 2.21h) Idem No. 5 9. 4PPPaR (Fig. 2.30a) P\\P\\kPa\\kR (Fig. 2.21i) Idem No. 4 10. 4PPPaR (Fig. 2.30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003590_kem.572.359-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003590_kem.572.359-Figure2-1.png",
        "caption": "Fig. 2 and 3.",
        "texts": [],
        "surrounding_texts": [
            "Geometric modeling and meshing.Table 1 shows the geometric parameters of a pair of meshing helical gears used in the analysis. The model established is as shown in Fig. 1. The finite element analysis uses full tooth model for transient analysis. When the model is discretized, all meshes use hexahedral mesh. The meshes of the tooth part should be refined as far as possible, and the meshes of the gear bodies are slightly larger. The discrete model is as shown in Application of load and boundary conditions. The gears in pair after discretization are linked via CONTACT_AUTOMATIC_SURFACE_TO_SURFACE. The analysis covers three conditions (see Table 2). The mechanical load is 1500r/min, the pinion is the driving gear, the thermal load is applied to the rings at 150 \u00b0 C, and the resistance moment is 600N.m. Complete modeling and submit for calculation. Note: \"+\" for apply and \"-\" for not apply."
        ]
    },
    {
        "image_filename": "designv11_100_0003464_amm.397-400.176-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003464_amm.397-400.176-Figure5-1.png",
        "caption": "Fig. 5 Bearing Pin, Planet Bearing and Planet Gear",
        "texts": [
            " The main carrier bearings fit over the collar and were held in place with two circlips. The collar assembly is shown in Fig. 4. The large spur gear was bolted to the carrier using three 8mm shoulder bolts. Shoulder bolts were used to ensure accurate location of the spur gear on to the carrier as any misalignment would affect its meshing with the pinion gear on the driving shaft. Three steel pins were press fit into recesses in the carrier. One of the two planet bearings on each gear fits over this pin and inside the machined hole in the planet gear as shown in Fig. 5. The other planet bearings fits into the opposite side of the planet gear with a steel bush located between the two bearings. The entire assembly is bolted together using an 8mm shoulder bolt. The accurately ground shoulder of the bolt fits tightly into the bearing bush, bearing pin and carrier ensuring that the planet gear is kept in the correct position relative to the Carrier. A cross section of the assembly is shown in Fig. 6. A small clearance on either side of each planet bearing allows some axial movement of the planet gear and reduces the sensitivity of the bearings to brinelling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure1-1.png",
        "caption": "Figure 1 \u2212 Classical shimmy model",
        "texts": [
            " The need of a deeper knowledge of the phenomenon, and in particular of the influence of some parameters, continues to be a priority for the setup and the control of the dynamic behavior of vehicles. Many mechanical models are available in the literature with different degree of approximations. The system stability is often deduced applying the Routh-Hurwitz criterion to the linearized equations of motion. In simple cases the stability condition may be analytically expressed; for example for the \u201cclassical shimmy model\u201d (Fig. 1), constituted by a rigid trailing wheel and a lumped lateral flexibility, in the hypothesis of no sideslip of the wheel, the stability condition expression is [1]: 2 Gab \u03c1> , \u03c1G being the system inertia gyrator and a, b the distances of the center of mass from the hinge point and from the center of the wheel, respectively. 1 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/75795/ on 04/09/2017 Terms of Use: http://www"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002783_1.3543587-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002783_1.3543587-Figure1-1.png",
        "caption": "Fig. 1. Double-spring torsion oscillator arrangement.",
        "texts": [
            " Both effective torsion constants and effective spring constants can be expressed in terms of adjustable parameters of the system. These expressions enable one to theoretically describe the motion of the hybrid oscillator and to calculate its period. A comparison of the translational and rotational interpretations teaches of their analogous mathematical properties and challenges the intuitive skills of those considering such systems. Torsion oscillator with two springs To construct the oscillator, two fixed springs are positioned to exert opposing torques on the disk (see Fig. 1). The adjustable parameters of the system include the force constants of the springs, the radii on which the springs exert restoring torques on the disk, and the moment of inertia of the disk. PASCO\u2019s introductory rotational apparatus3 is used for the rotational part of the oscillator, and two springs are attached horizontally by strings perpendicular to the radii on which they exert restoring forces. The fixed ends of the springs are attached to separate PASCO force sensors to record the force exerted by each spring"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003838_12.883154-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003838_12.883154-Figure1-1.png",
        "caption": "Figure 1: A schematic of the experimental setup used to measure blocked stresses of the liquid crystal networked that are stimulated by light.",
        "texts": [
            "aspx to above nematic-isotropic phase transition temperature and then subjected to 532 nm laser light over a time period ranging from 1 to 120 min for photocuring. More details on the synthesis and chemistry of the liquid crystal network are described elsewhere [16]. 3.1 Experimental Set-Up The specimens were cut to 15 mm\u00d71.84 mm\u00d715 \u03bcm and clamped at both ends with a small amount of pre-stretch. The pre-stretch is also varied to illustrate its effect on performance. One end of the specimen is connected to a load cell and another end is fixed. Figure 1 shows a schematic of the measurement setup. The load cell is manufactured by Interface (P/N: ULC-0.5N) and has a maximum capacity of 0.5 N. A light emitting diode (LED) (www.LEDSupply.com, P/N: Cree XPEROY-L1-0000-00901) is positioned about 1.5 cm away from the specimen and is utilized to induce light driven blocked stresses in the specimen. The royal blue LED has the wavelength ranging from 450 to 460 nm and provides 350 mW at 350 mA. Flashing frequencies of the LED are controlled by National Instruments and LabVIEW and the input voltage to the LED is controlled using a buckpuck (LEDdynamics, P/N: 3023-D-E-1000P)"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.15-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.15-1.png",
        "caption": "Fig. 7.15 The third and forth of four I-K solutions for a Stanford arm",
        "texts": [],
        "surrounding_texts": [
            "To further reduce the dimension of a combined isometric embedding given in (7.55), let us now pay more attention to the last link, i.e., the kinematically \u201cbusiest\u201d link of an n-joint open serial-chain robot. Obviously, the isometric embedding of link n, zn = \u03b6n(q) can embed the C-manifold Mn of link n into R 9 by (7.54) with i = n. Intuitively, because any joint movement will contribute a non-trivial part to the last link motion, the dimensions of Mn for the last link and Mc for the entire robot should both be equal to n. Now the question is can the mapping zn = \u03b6n(q) of the last link also embed the Cmanifold Mc of the entire robot into R 9? If so, we would achieve a significant model reduction. Let zn|Mn and z|Mc be, respectively, a point on the last link C-manifold Mn sitting in R 9 and a point on the entire robot C-manifold Mc in R 9n. If we define a common local coordinate system by using the n robotic joint positions q = (q1 \u00b7 \u00b7 \u00b7 qn)T on both Mn and Mc, then we seek a continuous one-to-one in both directions between Mn and Mc via the common local coordinates q, i.e., zn|Mn \u2190\u2192 q|Z \u2190\u2192 z|Mc . We can directly observe that through the common q under the quotient operation by the integer group Z for every joint angle, each point on Mn determined by zn = \u03b6n(q) = Zn 0 (q)\u03ben can respond to a unique point on Mc determined by z = \u03b6(q) = Z(q)\u03be, and vice verso, as long as the mapping \u03b6n : Mn \u2192 R 9 of the last link is a smooth embedding under the robotic joint positions in q as the local coordinate system. This reaches the following lemma: Lemma 3. For an open serial-chain robotic system with n joints, let the n joint positions in q = (q1 \u00b7 \u00b7 \u00b7 qn)T be defined as a common n-tuple local coordinate system on both the C-manifold Mn of the last link and C-manifold Mc of the entire robot. If (7.54) with i = n can be a smooth embedding that sends Mn into R 9, then Mn is diffeomorphic to Mc, i.e., Mn Mc. Note that if the number of joints of the robot arm is n \u2264 6 in the above lemma, it may have more chances for the mapping \u03b6n to be a smooth embedding. If n > 6, such an embedding condition on Mn would not be possible. In fact, n > 6 is a redundant robot case, as discussed in Chapter 5, and every redundant robotic system has a continuum of multi-configurations. In the redundant robot case, the combined C-manifold Mc of the entire robot may still be 6-dimensional and so is the C-manifold Mn of the last link, unless additional subtasks are augmented to the 6 d.o.f. rigid motion as a redundant kinematic output. In addition, Lemma 3 reveals a differentiable (smooth) topological equivalence between the two C-manifolds Mn Mc under the condition that (7.54) with i = n is a smooth embedding if all the joint positions of the robot constitute a local coordinate system. A number of non-redundant robot arms can directly satisfy the above condition due to their special mechanical structures. A robot having more independent prismatic joints may have more chances to satisfy the smooth embedding condition. If, in general, the smooth embedding condition cannot hold by the last link mapping alone, we have to augment the Euclidean zspace by adding one or more links until the augmented mapping is qualified to be a smooth embedding. Obviously, if the final augmentation covers all n links of the robot, it must be smoothly embeddable, because the total Riemannian metric W = JT J that is also the inertial matrix of the robot should always be non-singular. However, in most cases, only a small number of links need to be selected, say 1 < k < n, which, of course, include the last link to form an augmented mapping into R 9k to meet the smooth embeddability condition. Such an augmented smooth embedding \u03b6nk with the smallest k is referred to as the minimum embedding of the C-manifoldMc for the robotic system. Therefore, based on the principle of Lemma 3, under the smallest number k, it is always true that Mnk Mc for every open serial-chain robotic system, even for a redundant robot. Now, the new question is how to find such a minimum embedding? Recalling Definition 6, a mapping between two manifolds can be an embedding if it is a one-to-one immersion. Thus, by intuition, this one-to-one condition holds if no more multi-configuration case can be found among the remaining n\u2212 k links (k < n) whenever the selected k links are stationary. It is well known in robotic kinematics that a 6-joint PUMA robot with a shoulder offset (i.e., d2 = 0) has up to eight different configurations (i.e., eight different sets of I-K solutions) if the last link (link 6) is fixed. Whereas for the 6-joint PUMA-like arm without the shoulder offset (d2 = 0), the number of different configurations is down to four. In contrast, a 6-joint Stanford robot with a shoulder offset (d2 = 0) has in total four different configurations whenever link 6 is motionless. These four different sets of the I-K solutions for the Stanford arm have been animated in MATLABTM , as realistically depicted in Figures 7.14 and 7.15. Therefore, to only pick the last link of the 6-joint Stanford arm (or PUMA arm) is insufficient to constitute a smooth embedding for the C-manifold Mc if all the 6 joint positions are defined as the local coordinates on the last link C-manifold M6. However, it is also observable that if we further select link 1 and link 4, in addition to link 6 for the Stanford arm to augment a relatively larger mapping for Mc, then it becomes embeddable because all the possible multi-configurations disappear whenever the selected three links are stationary. In this case, the minimum embedding \u03b6nk is determined by augmenting the mappings (7.54) of link 6, link 4 and link 1, and thus k = 3. The total dimension of this minimum embedding is now 9k = 9 \u00d7 3 = 27 that is much less than 9n = 9\u00d7 6 = 54 in the regular augmentation (7.55), and just touches the dimension predicted by the Nash-Greene Theorem for isometrically embedding the C-manifold Mc of the Stanford arm. Furthermore, if each link can adopt the two-point model, i.e., only one radial vector r1c for each link with a diagonal inertia tensor, then the dimension of the minimum embedding will be further reduced to 6k = 18. Actually, to test whether a mapping \u03b6 : Mn \u2192 R m for an n-dimensional C-manifold Mn is an embedding, we may just check its Jacobian matrix J = \u2202\u03b6/\u2202q to see if rank(J) = n, the full rank at every point on Mn. If the Jacobian matrix of a mapping has full rank, then every multi-configuration case will disappear. Therefore, to find the minimum embedding of the Cmanifold Mc for a robot by the above k-augmentation procedure, testing its Jacobian matrix J to see if it is full-ranked is the easiest way to forecast the result. Moreover, two smoothly topologically equivalent (diffeomorphic) Riemannian manifolds in a common ambient Euclidean space can always be smoothly deformed to make them isometric to each other. Even though their ambient Euclidean spaces are in different dimensions, it may still be possible, but not always. We refer to this smooth deformation process as an isometrization [23]. Since the Riemannian metric of any smooth C-manifold is given by W = JTJ , where J is the Jacobian matrix of the C-manifold, based on (7.55) and (7.56), each component of W can be written as wi j = gTi gj = \u03beT \u2202Z \u2202qi T \u2202Z \u2202qj \u03be = tr ( \u2202ZT \u2202qi \u2202Z \u2202qj \u03a8 ) , for i, j = 1, \u00b7 \u00b7 \u00b7 , n, (7.57) where \u03a8 = \u03be\u03beT is called a parameter matrix and tr(\u00b7) is the trace of a square matrix (\u00b7). We now give a formal definition regarding the isometrization below: Definition 8. Let two n-dimensional C-manifolds Mn a and Mn b be smoothly embedded into R ma and R mb , respectively, on both of which a common local coordinate system {q1, \u00b7 \u00b7 \u00b7 , qn} is defined. Then,Mn a is said to be isometrizable to Mn b if for each element (wa) i j of the metric Wa on Mn a , i.e., (wa) i j = \u03beTa \u2202ZT a \u2202qi \u2202Za \u2202qj \u03bea = tr ( \u2202ZT a \u2202qi \u2202Za \u2202qj \u03a8a ) , there is a real parameter function D(\u00b7) such that the new parameter matrix \u03a8 \u2032 a = D(\u03a8b) can make the following equation hold globally: (w\u2032 a) i j = tr ( \u2202ZT a \u2202qi \u2202Za \u2202qj \u03a8 \u2032 a ) = tr ( \u2202ZT b \u2202qi \u2202Zb \u2202qj \u03a8b ) = (wb) i j for each i, j = 1, \u00b7 \u00b7 \u00b7 , n. We call this real parameter function D(\u00b7) a deformer. Therefore, to respond to the model reduction challenge, the most promising approach is to expect the minimum embedding \u03b6nk of the C-manifold Mnk to be isometrizable to the combined C-manifold Mc through the above deformation with a certain deformer. However, Mc has been isometrically embedded into R 9n, and may possess more geometrical details than Mnk that is embedded into a smaller Euclidean space R9k with k < n. In other words, R9k may not be spatial enough to allowMnk to be isometrizable toMc even though topologically Mnk Mc. Nevertheless, we are dealing with a special open serial-chain robot case. As commonly experienced, when trying to derive a symbolical form of kinetic energyKi for link i of an n-joint serial-chain robotic system with (1 < i \u2264 n), we must take into account every velocity of both translation and rotation of link i \u2212 1, in addition to all the terms contributed by the motion of link i itself, because the latter is often imposed by the former. Therefore, the sum Knk of all the kinetic energies Ki\u2019s for the k links involved in the C-manifold Mnk that is associated with the minimum embedding \u03b6nk should be able to cover every factor contained in the total kinematic energy K = \u2211n i=1Ki of the robot. In other words, we can always adjust each dynamic parameter in Knk to make Knk equal to K. Because the coefficient of each q\u0307iq\u0307j in K given by (7.47) is just the metric element wij , this parameter adjustment is virtually a smooth deformation mentioned in Definition 8. Hence, we reach the following theorem: Theorem 4. For an n-joint open serial-chain robot dynamic system, its minimum embeddable C-manifold Mnk is diffeomorphic to the combined Cmanifold Mc of the entire robotic system. If all the robotic joints are revolute, then, Mnk is also isometrizable to Mc. This theorem underlies a fundamental principle of robot dynamic model reduction. Namely, the lower-bound of dynamic model reduction in the sense of topology is a subsystem with the minimum embeddable C-manifold for every open serial-chain robotic system. In other words, the robot dynamic model cannot be further reduced below the lower-bound, otherwise both the embeddability and isometrizability will no longer be guaranteed, and may even cause a catastrophe of topological structure destruction [38, 39]. Theorem 4 indicates that if every joint is revolute for a robotic system, then the minimum embeddable C-manifold Mnk is not only diffeomorphic to the combined C-manifold Mc, but is also isometrizable to Mc. If one of the joints is prismatic, Mnk is still diffeomorphic to Mc, but Mnk may have to further include the prismatic link in order to meet the isometrizability. The reason is that since the variable for a prismatic joint is a length di, instead of an angle, as one of the local coordinates, it will often be mixed with the dynamic parameters to be deformed together in the Riemannian metrics {wij}, causing unavailability of an effective deformer to distinguish the joint variable di from the dynamic parameters. As a first example, let us look at a well-known inverted pendulum system that consists of only two links: a linearly moving cart and a rotating pole mounted on the cart, as shown in Figure 7.16. With the joint positions x1 and \u03b82 defined as two local coordinates q = (x1 \u03b82) T for its 2-dimensional combined C-manifold, we can readily find its inertial matrixW by extracting all the coefficients of q\u0307iq\u0307j terms for i, j = 1, 2 from its kinetic energy K, i.e., W = ( m1 +m2 \u2212m2lc2s2 \u2212m2lc2s2 m2l 2 c2 + I2 ) , (7.58) where m1 and m2 are the masses of the cart and the top pole as link 1 and link 2, respectively, lc2 is the length between the revolute joint axis and the mass center of the pole, as link 2, and I2 is the moment of inertia of link 2. It can be directly seen that if link 2 is fixed, the cart (link 1) will have no chance to move. Thus, link 2 determines the minimum embeddable Cmanifold Mnk of the system with n = 2 and k = 1. For this planar system, the position vector of the mass center of link 2 with respect to the base is pc20 = ( ax1 + bc2 bs2 ) , where a and b are two dynamic parameters of link 2. Since this pole rotates only about one single axis, we can simply use h\u03b82 with a new parameter h to represent the rotation of link 2 under 0 \u2264 \u03b82 < 2\u03c0. Therefore, the smooth embedding for Mnk can be written as \u23a7\u23a8 \u23a9 z12 = ax1 + bc2 z22 = bs2 z32 = h\u03b82. (7.59) The Jacobian matrix of this minimum embedding z2 = \u03b62(q) becomes J2 = \u2202\u03b6 \u2202q = \u239b \u239d a \u2212bs2 0 bc2 0 h \u239e \u23a0 , the rank of which is obviously rank(J) = 2, as long as both a = 0 and h = 0. Its Riemannian metric turns out to be W2 = JT 2 J2 = ( a2 \u2212abs2 \u2212abs2 b2 + h2 ) . Comparing this with the inertial matrix W in (7.58), we can see that if a = \u221a m1 +m2, b = m2lc2\u221a m1 +m2 and h = \u221a m1m2 m1 +m2 l2c2 + I2, (7.60) thenW2 =W . This demonstrates that the minimum embeddable C-manifold given by (7.59) is isometrizable to the combined C-manifold Mc of the inverted pendulum system, and equation (7.60) is just its deformerD(\u03a8) during the isometrization process. Figure 7.17 visualizes the minimum embeddable C-manifold M21 in R 3 based on (7.59) when a = 4, b = 1.5 and h = 1 in MATLABTM . This 2- surface is diffeomorphic to the compact cylindrical surface S1 \u00d7 I1 [28, 29] after \u03b82 in z32 is confined within [0, 2\u03c0) by a quotient operation R 1/Z, where Z is the integer additive group. In other words, once the z32 component of the 2-surface reaches 2\u03c0, it should be imagined to return immediately back to the zero and start over again. The second example is a three-revolute-joint (RRR-type) planar arm, as shown in Figure 7.18. According to equation (7.54), the last link i = n = 3 has a 3-dimensional C-manifold M3 that can be sent to R 3 by \u23a7\u23a8 \u23a9 z13 = ac1 + bc12 + dc123 z23 = as1 + bs12 + ds123 z33 = h(\u03b81 + \u03b82 + \u03b83) (7.61) with four non-zero parameters a, b, d and h, where sij = sin(\u03b8i + \u03b8j), cij = cos(\u03b8i + \u03b8j), sijk = sin(\u03b8i + \u03b8j + \u03b8k) and cijk = cos(\u03b8i + \u03b8j + \u03b8k) for i, j, k = 1, 2, 3. Its Jacobian matrix becomes J3 = \u2202\u03b63 \u2202q = \u239b \u239d \u2212as1 \u2212 bs12 \u2212 ds123 \u2212bs12 \u2212 ds123 \u2212ds123 ac1 + bc12 + dc123 bc12 + dc123 dc123 h h h \u239e \u23a0 . (7.62) The determinant of this 3 by 3 Jacobian matrix is calculated as det J3 = abhs2 = abh sin \u03b82. This shows that M3 is singular at \u03b82 = 0 or \u00b1\u03c0. In fact, we can see that for each fixed z3 = (z13 z 2 3 z 3 3) T \u2208 R 3, \u03b81 through \u03b83 can have two different sets of I-K solution, as long as they keep the sum \u03b81 + \u03b82 + \u03b83 fixed while one uses \u03b82 and the other one takes \u2212\u03b82, as shown in Figure 7.18. This is a typical multi-configuration phenomenon. We may thus construct a minimum embeddable C-manifold by adding link 1, which will clearly make the above multi-configuration disappear. Since link 1 has only a single-axis rotation, we can augment (7.61) by z4 = h1\u03b81 so that a possible minimum embedding would be \u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 z1 = ac1 + bc12 + dc123 z2 = as1 + bs12 + ds123 z3 = h(\u03b81 + \u03b82 + \u03b83) z4 = h1\u03b81 (7.63) with a new non-zero parameter h1. It can be easily verified that the rank of the following new 4 by 3 Jacobian matrix J is always full: J = \u2202\u03b632 \u2202q = \u239b \u239c\u239d \u2212as1 \u2212 bs12 \u2212 ds123 \u2212bs12 \u2212 ds123 \u2212ds123 ac1 + bc12 + dc123 bc12 + dc123 dc123 h h h h1 0 0 \u239e \u239f\u23a0 , and the set of all four equations in (7.63) is also one-to-one. Thus, it shows that equation (7.63) is a minimum embedding with n = 3 and k = 2. The Riemannian metricW32 on the minimum embeddable C-manifold can be found by W32 = JTJ = \u239b \u239d w11 w12 w13 w12 b2 + 2bdc3 + d2 + h2 bdc3 + d2 + h2 w13 bdc3 + d2 + h2 d2 + h2 \u239e \u23a0 , where w11 = a2 + b2 + 2abc2 + 2adc23 + 2bdc3 + d2 + h2 + h21, and w12 = b2 + abc2 + adc23 + 2bdc3 + d2 + h2, w13 = adc23 + bdc3 + d2 + h2. One can also verify without difficulty that by adjusting all the parameters a, b, d, h and h1, the metric W32 = JTJ can be equal to the inertial matrix W of this RRR-type planar robot. Such a parameter adjustment just plays the role of deformer D(\u03a8) in the C-manifold isometrization. In contrast to the above RRR-type planar robot arm, let us revisit the RPR-type planar robot, as given in Figure 7.7 or 7.9. At first glance, it seems sufficient enough to pick up only the last link to represent the minimum embeddable C-manifold for this particular robot, because the configuration becomes unique whenever the last link is motionless. If so in this case, n = 3 and k = 1. Based on the previous kinematics analysis of this planar robot, the tip position vector and the position vector of frame 2 can be found as pt0 = \u239b \u239d d2s1 + d4s13 \u2212d2c1 \u2212 d4c13 0 \u239e \u23a0 , and p20 = \u239b \u239d d2s1 \u2212d2c1 0 \u239e \u23a0 where d4 is the length of the last link. Since the z-component of pt0 is zero due to the planar arm, while the number of joints is three, we must replace this zero by the orientation of the last link, which can be uniquely determined by \u03b81 + \u03b83. Thus, the proposed embedding is given as follows: \u23a7\u23a8 \u23a9 z13 = ad2s1 + bs13 z23 = \u2212ad2c1 \u2212 bc13 z33 = h(\u03b81 + \u03b83) (7.64) with parameters a, b and h to be adjusted. Then, the Jacobian matrix becomes J3 = \u2202\u03b63 \u2202q = \u239b \u239d ad2c1 + bc13 as1 bc13 ad2s1 + bs13 \u2212ac1 bs13 h 0 h \u239e \u23a0 . Its determinant is det(J) = \u2212a2hd2 and is never zero as long as the prismatic joint value d2 = 0. Now, the Riemannian metric for this minimum embeddable C-manifold M31 turns out to be W31 = JT 3 J3 = \u239b \u239d a2d2 + b2 + h2 + 2abd2c3 \u2212abs3 abd2c3 + b2 + h2 \u2212abs3 a2 \u2212abs3 abd2c3 + b2 + h2 \u2212abs3 b2 + h2 \u239e \u23a0 . Comparing this with the inertial matrixW of this RPR arm derived in equation (7.23), we find that if a2 = m2 +m3, ab = m3lc3, and b2 + h2 = I3 +m3l 2 c3, every element wij in W can be matched by W31 except the first one w11 that contains a mixed term m2(d2 \u2212 lc2)2 between the prismatic joint variable d2 and the dynamic parameters. Therefore, while equation (7.64) is an embedding, it has not been isometrizable yet to exactly match the inertial matrix W . This phenomenon reveals a special complication due to the prismatic joint in isometrization. By further augmenting the second prismatic link, the new embedding becomes: \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 z13 = a1d2s1 + b1s13 z23 = \u2212a1d2c1 \u2212 b1c13 z33 = h1(\u03b81 + \u03b83) z43 = a2(d2 \u2212 b2)s1 z53 = \u2212a2(d2 \u2212 b2)c1 z63 = h2\u03b81 (7.65) along with six parameters a1, b1, h1, a2, b2 and h2 to be adjusted, where d2 \u2212 b2 indicates a distance to converge to the mass center of link 2. The Jacobian matrix of the new proposed embedding (7.65) can be derived as J3 = \u2202\u03b63 \u2202q = \u239b \u239c\u239c\u239c\u239c\u239c\u239d a1d2c1 + b1c13 a1s1 b1c13 a1d2s1 + b1s13 \u2212a1c1 b1s13 h1 0 h1 a2(d2 \u2212 b2)c1 a2s1 0 a2(d2 \u2212 b2)s1 \u2212a2c1 0 h2 0 0 \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 . Thus, the Riemannian metric becomes W32 = JT 3 J3 = \u239b \u239d w11 \u2212a1b1s3 a1b1d2c3 + b21 + h21 \u2212a1b1s3 a21 + a22 \u2212a1b1s3 a1b1d2c3 + b21 + h21 \u2212a1b1s3 b21 + h21 \u239e \u23a0 , where w11 = a21d 2 2 + b21 + h21 + 2a1b1d2c3 + a22(d2 \u2212 b2)2 + h22. In comparison with equation (7.23) again, the new Riemannian metricW32 has now sufficient factors to match the inertial matrix W . Namely, let a1 = \u221a m3, a2 = \u221a m2, b1 = \u221a m3lc3, h1 = \u221a I3, and b2 = lc2, h2 = \u221a I1 + I2 +m1l2c1. Then every element wij inside the new metric W32 can exactly match with the inertial matrix W in (7.23). The last example is to dynamically model a general fully parallel-chain mechanism, such as a 6-6 Stewart platform, as studied in [36, 37]. If we focus on the top mobile disc, according to equation (7.25), its kinetic energy can be written as follows: K = 1 2 mv6T0 v60 +mv6T0 CT 6 \u03c9 6 0 + 1 2 \u03c96T 0 \u03936\u03c9 6 0, where m and \u03936 are the mass and inertia tensor with respect to frame 6 on the top disc, respectively, and C6 is the skew-symmetric matrix of the mass center coordinates with respect to frame 6. If the origin of frame 6 is defined at the mass center of the top mobile disc, the second term of the kinetic energy K vanishes and \u03936 = \u0393c. Because the top disc is a single rigid body, Theorem 3 can be directly applied to it. Then, an isometric embedding that can send the 6-dimensional C-manifold M6 6 of the top disc motion to Euclidean 9-space R 9 is given by z = \u03b6(q) = Z\u03be = \u239b \u239d R6 0 p60 O R6 0 O R6 0 \u239e \u23a0 \u03be (7.66) with a 10 by 1 dynamic parameter vector \u03be based on equation (7.54). In order to find the Riemannian metric W = JT J endowed on the CmanifoldM6 for the top mobile disc of the 6-6 Stewart platform, it is required to know a local coordinate system q to be defined on M6. If we pick the 6 prismatic joint lengths qi = li\u2019s to form the local coordinate system, then intuitively, the top disc will have no multi-configuration chance so that the mapping (7.66) is a minimum embedding for the C-manifold Mc of the entire Stewart platform. The Jacobian matrix of the minimum embedding should be J = \u2202\u03b6/\u2202q. However, this turns back to the forward kinematics (F-K) problem. Namely, it is difficult or even impossible to find an explicit closed form for either R6 0 or p60 as a function of q = (q1 \u00b7 \u00b7 \u00b7 q6)T = (l1 \u00b7 \u00b7 \u00b7 l6)T before taking derivatives to determine J and then W . If we define a local coordinate system alternatively other than the 6 piston lengths, it may be possible to reach a Riemannian metric result. For instance, let the 6 local coordinates be defined by the variables in Cartesian space: qc = \u239b \u239c\u239c\u239c\u239c\u239c\u239d x y z \u03c6 \u03b8 \u03c8 \u239e \u239f\u239f\u239f\u239f\u239f\u23a0 = ( p60 \u03c160 ) , (7.67) where x, y and z are the three coordinate components of p60, and \u03c6, \u03b8 and \u03c8 are the roll, pitch and yaw Euler angles to represent R6 0, see equation (3.7) in Chapter 3. Namely, p60 = \u239b \u239d x y z \u239e \u23a0 , R6 0 = \u239b \u239d c\u03c6c\u03b8 \u2212s\u03c6c\u03c8 + c\u03c6s\u03b8s\u03c8 s\u03c6s\u03c8 + c\u03c6s\u03b8c\u03c8 s\u03c6c\u03b8 c\u03c6c\u03c8 + s\u03c6s\u03b8s\u03c8 \u2212c\u03c6s\u03c8 + s\u03c6s\u03b8c\u03c8 \u2212s\u03b8 c\u03b8s\u03c8 c\u03b8c\u03c8 \u239e \u23a0 . Then, the 9 by 1 column of the embedding z = \u03b6(qc) = Z\u03be in (7.66) after the structure matrix Z is linearly combined by the dynamic parameters in \u03be can be expressed explicitly in terms of the above 6 new local coordinates. The Jacobian matrix J = \u2202\u03b6/\u2202qc as well as the Riemannian metric W = JT J can also be determined successfully. However, under the new local coordinate system (7.67), due to the nonuniqueness between the roll, pitch and yaw Euler angles and the rotation matrix, see the example in Chapter 3, there will be at least one multiconfiguration case. In other words, the mapping z = \u03b6(qc) will no longer be a minimum embedding. Even if it could be remedied by imposing a constraint on the three Euler angles, once the Riemannian metric W on the Cmanifold is determined, by substituting it into the Lagrange equation (7.15), the control input \u03c4c will no longer be the desired six piston forces. Instead, it should be a some Cartesian force vector in correspondence to the new local coordinates defined in Cartesian space and given by (7.67), which requires a conversion. Naturally, we may take advantage of the statics equation (5.30) from Chapter 5 for the Stewart platform to convert the control input \u03c4c that is resolved by the Lagrange equation in Cartesian space to a piston joint force vector by \u03c4 = J\u22121 0 F0. To do so, according to equation (3.12) in Chapter 3, the angular velocity of the top disc is \u03c96 0 = \u239b \u239d 0 \u2212s\u03c6 c\u03c6c\u03b8 0 c\u03c6 s\u03c6c\u03b8 1 0 \u2212s\u03b8 \u239e \u23a0 \u239b \u239d \u03c6\u0307 \u03b8\u0307 \u03c8\u0307 \u239e \u23a0 = D\u03c1\u030760. Hence, the 6 by 1 Cartesian velocity vector becomes V0 = ( v60 \u03c96 0 ) = ( I O O D )( p\u030760 \u03c1\u030760 ) = Bq\u0307c, where the 6 by 6 coefficient matrix is denoted by B, and I and O are the 3 by 3 identity and zero matrix, respectively. Based on the Jacobian equation (5.29) in Chapter 5 for the Stewart platform, q\u0307 = JT 0 V0 = JT 0 Bq\u0307 c. (7.68) This arrives at a kinematic conversion in tangent space between the prismatic joint position vector q = (l1 \u00b7 \u00b7 \u00b7 l6)T and the local coordinate system qc defined by (7.67). Furthermore, the mechanical power seen in joint space is P = q\u0307T \u03c4 , where \u03c4 = (f1 \u00b7 \u00b7 \u00b7 f6)T is the piston actuating force vector, while the same power seen in Cartesian space is P = q\u0307cT \u03c4c for a Cartesian force \u03c4c that is corresponding to the local coordinates in (7.67). Based on the principle of energy conservation, q\u0307T \u03c4 = q\u0307cT \u03c4c. Substituting (7.68) into here, we have q\u0307cTBTJ0\u03c4 = q\u0307cT \u03c4c so that \u03c4c = BTJ0 \u03c4. (7.69) This new statics equation is similar to (5.30) in Chapter 5. Thus, if a control law \u03c4c can be resolved through the Lagrange equation (7.15), then the piston actuating forces can be determined by the new statics equation (7.69), i.e., \u03c4 = J\u22121 0 B\u2212T \u03c4c, provided that both the Jacobian matrix J0 and the matrixB are non-singular. Therefore, the conversion issue between \u03c4c and \u03c4 has been solved by the statics in (7.69), but the one-to-one, or multi-configuration issue still remains unfixed as a barrier of achieving the minimum isometrizable embedding. In conclusion, if we insist in using the 6 piston lengths to define a local coordinate system on the C-manifoldMc of the Stewart platform, we will not be able to continue its Riemannian metric determination towards a successful adaptive control design until the F-K problem for the closed parallel-chain systems is resolved. Finally, it should be pointed out that all the parameter adjustments in the above examples are just to illustrate the possibility of isometrization, and are not required to do so by hand at all. Instead, the computer program will automatically adjust the model parameters towards the isometrization if an effective adaptive control algorithm is implemented. The detailed introduction and discussion on adaptive control as well as a 3D robotic system example will be given in the next chapter."
        ]
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.31-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.31-1.png",
        "caption": "Fig. 2.31 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4PPaRP (a) and 4PPaPR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 18, limb topology P||Pa\\R||P (a) and P||Pa\\P|| R (b)",
        "texts": [
            "21d) The last revolute joints of the four limbs have parallel axes 5. 4PRPPa (Fig. 2.28a) P||R\\P||Pa (Fig. 2.21e) The second joints of the four limbs have parallel axes 6. 4PPRPa (Fig. 2.28b) P\\P||R\\Pa (Fig. 2.21f) The third joints of the four limbs have parallel axes 7. 4PRPPa (Fig. 2.29a) P\\R||P\\Pa (Fig. 2.21g) Idem No. 5 8. 4PRPaP (Fig. 2.29b) P\\R\\Pa \\kP (Fig. 2.21h) Idem No. 5 9. 4PPPaR (Fig. 2.30a) P\\P\\kPa\\kR (Fig. 2.21i) Idem No. 4 10. 4PPPaR (Fig. 2.30b) P\\P||Pa\\R (Fig. 2.21j) Idem No. 4 11. 4PPaRP (Fig. 2.31a) P||Pa\\R||P (Fig. 2.21k) The before last joints of the four limbs have parallel axes 12. 4PPaPR (Fig. 2.31b) P||Pa\\P|| R (Fig. 2.21l) Idem No. 4 13. 4PPaPR (Fig. 2.32a) P\\Pa||P\\\\R (Fig. 2.21m) Idem No. 4 14. 4PPaRP (Fig. 2.32b) P\\Pa\\kR\\kP (Fig. 2.21n) Idem No. 4 15. 4RPPaP (Fig. 2.33a) R\\P||Pa\\kP (Fig. 2.21o) The first revolute joints of the four limbs have parallel axes 16. 4RPPPa (Fig. 2.33b) R||P\\P||Pa (Fig. 2.21p) Idem No. 15 17. 4PaPPR (Fig. 2.34a) Pa||P\\P||R (Fig. 2.21q) Idem No. 4 18. 4PaPPR (Fig. 2.34b) Pa\\P\\kP\\kR (Fig. 2.21r) Idem No. 4 19. 4PaRPP (Fig. 2.35a) Pa\\R||P\\P (Fig. 2.21s) Idem No. 11 20"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002230_icara.2011.6144893-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002230_icara.2011.6144893-Figure5-1.png",
        "caption": "Figure 5. 4DoF Model",
        "texts": [
            " White areas in Fig. 4(a) are hard areas to which outside force can be applied. The only hard areas around the shoulder are around the articulatio acromioclavicularis and around the scapulothoracic joint. These are the points to which the trainer will apply guidance force to let the trainee learn the proper motion of the scapula and the shoulder joints. Fig. 4(b) shows corresponding points in the 4DoF model. III. SCAPULA MOTION AND SCAPULA MOTION ANALYSIS A. Aanlysis Method by the 4DoF Model In Fig. 5, \u2211 0 is the coordinate system on the human body, and \u2211 p is the coordinate system on the scapula ( 4321 CCCC ) with its origin at p O . Position and pose of the scapula in the 4DoF model is expressed by four angles ( 1 \u03b8 , 2 \u03b8 , 3 \u03b8 , 4 \u03b8 ). 1 R , 2 R , 3 R , 4 R represent length variables of four sliding actuators. The angle 1 \u03b8 is the angle of 113 BAA . The angle 2 \u03b8 is the angle between the triangle 113 BAA and the XY plane of \u2211 0 . The angle 3 \u03b8 is the angle between the triangle 113 BAA and the triangle 331 BAB "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003396_amm.275-277.174-Figure6-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003396_amm.275-277.174-Figure6-1.png",
        "caption": "Fig. 6 The meshes of quarter model of crankshaft section",
        "texts": [
            "2 mm displacement is applied on the upper plane of half roller model in the direction of feed. In the second load step, the displacement in the direction of feed is kept invariable, and 15\u00baangular displacement is applied on the upper plane of half roller model to simulate the relative rotation between roller and stepped shaft. In the last load step, -0.2 mm displacement is applied on the upper plane of half roller model to unload the roller. The bottom plane of half stepped shaft is restrained all the time. The mesh model is created in Ansys-ICEM as shown in Figure 6. The quadrilateral elements in the symmetry plane of half roller and half stepped shaft model are generated manually, and finer mesh is adopted near contact position. Then the hexahedral solid elements are created by rotating the plane elements about their axes respectively. The density of element between plane A and C is much higher than other regions. The element size near rolling location is about 0.05mm. Results and Discussion. After the roller is unloaded, the stresses in plane B could be regarded as residual stress which induced by fillet rolling process",
            " The most significant stress component for the fatigue analysis is the normal stress \u03c3\u03b8 in the polar coordinate system near fillet as shown in Fig. 4, which could be marked as \u03c3\u03b8r. The distributions of \u03c3\u03b8r in different depth a along \u03b8 are shown in Fig. 5. Boundary Conditions and FE Model. The geometry parameters of the crankshaft section not shown in Fig. 1 are list in Table 3. Because the geometry and bending loads are symmetrical, quarter model of crankshaft section is selected. The geometry model is divided into some blocks, and then hexahedron meshes are generated as shown in Fig. 6. In the fillet region of main journal and rod journal where stresses concentrate, local mesh refinement approach is used. Symmetry constraints are set in the symmetry plane of model, and displacements in vertical direction are restrained by weak spring. After that, static bending moment of 2400 Nm is applied to the model. All the pre process operations are completed in HyperMesh. Results and Discussion. Under cycle bending loads, the amplitude of normal strain component \u03b5\u03b8 in dangerous section ,which could be marked as \u03b5\u03b8b, is closely related to the life of crankshaft, so it is always used in local stress-strain approach"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002567_icmtma.2011.343-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002567_icmtma.2011.343-Figure3-1.png",
        "caption": "Figure 3. Sketch of a wire race ball bearing",
        "texts": [
            " HERTZ CONTACT THEORY-BASED CONTACT STRESS Hertz contact stress formula is widely used in engineering applications, such as simply evaluating the stress and deformation on spherical pair, ball bearings, elastic hydrodynamic lubrication (EHL) of gears, etc. Hertz point contact problem is related to the terms of two arbitrary elastic curved body contacting only at one point in the initial state. In the loading, deformation and stress of the contact area have a local nature, they will decay quickly with increasing distance from the point of contact[4]. The contact area is of ellipse whose size parameters are described by the Hertz contact stress formula. The contact model of two balls is shown in Figure 3. The contact between balls and wires can be seen as Hertz contact problems in the case of ignoring friction[1]. By Hertz contact theory, the expression of contact deformation and contact stress are[4]: 1/33 ( ) 2 E Q a A 1/33 ( ) 2 E Q b B 2/33 ( ) 2 2E Q K 1/22 2 max( , ) 1 x y x y a b Integrate surface pressure distribution along the contact region is equal to the total load, then: ( , )Q x y d Integral region is the contact area between the two objects. Based on the above formula, get: max 3 2 Q ab Introduce the main curvature sum and main curvature function F ( ): 11 12 21 22 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure24-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure24-1.png",
        "caption": "Fig. 24 Tip chamfering with EM tip end\u2014spiral bevel pinion",
        "texts": [
            " And again, End Mills can be used to chamfer some tooth edges, but now the options are more limited because the spiral angle can become quite significant at the Heel tooth edge for large offset hyoid pinions. For example, consider tip chamfering with the cutting edge of the End Mill as shown in Fig. 23. While chamfering the convex flank causes no issue (left, Fig. 23), chamfering the concave flank (right, Fig. 23) causes a collision risk between the tool spindle and the turn table. This approach is therefore not acceptable and the EM tool tip is to be used which, as shown in Fig. 24, causes no issue on either flank. Figure 25 shows the Tip edges of a small spiral bevel pinion chamfered using an EM. The chamfers are almost small enough to confuse with deburring. When chamfering tooth Toe and Heel edges, other considerations arise. For one, when chamfering the Toe of a pinion with a small pitch cone angle, again the turn table angle is likely to exceed the machine\u2019s limit and collide with the tool spindle (left, Fig. 26). Beyond this, even a small diameter EndMill is likely to interfere with the concave tooth flank when chamfering the bottom of the tooth (right, Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-Figure16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-Figure16-1.png",
        "caption": "Figure 16. Included Angles between Planet Gears",
        "texts": [
            " Equally spaced planets (Eqn. (1)) will be symmetric, but symmetric planets (Eqn. (2)) are not necessarily equally spaced. The condition for symmetric planet gears is: m Integer N NN SR 1)( \u00b1= + (2) m=1,2,3, ..., N/2, N = even number m=1, N = odd number Figure 15. Arrangements of Six Planet Gears The included angle between any two planet gears is simply 360\u25cb / N, if the planet gears are equally spaced. On the other hand, if the planets are not equally spaced (asymmetric arrangement) as shown on the right side of Figure 16, the included angle becomes: \u239f \u23a0 \u239e \u239c \u239d \u239b + \u22c5 + = N NNiround NN SR SR i )(*360o\u03b8 i=1...N (3) For maximum load carrying capacity, the goal is to include as many planet gears in the available annular space as possible. Asymmetric arrangements will waste the precious space, so it is not recommended for compound planetary GRAs. The maximum number of planets that can be assembled in a symmetric arrangement is governed by: N CD ODP > \u23a5\u23a6 \u23a4 \u23a2\u23a3 \u23a1\u2212 *2 sin 1 \u03c0 (4) where ODp is the outside diameter of planet gears, and CD is the center distance"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003431_2013-36-0272-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003431_2013-36-0272-Figure5-1.png",
        "caption": "Figure 5 \u2013 Absolute and relative movements",
        "texts": [
            " Let us take a segment AB of the tire belt, as shown in Figure 4, with length l0 and mass m, resulting in a linear mass density given by: (2) If this segment is deformed (stretched) by some force F, its new length will be l and the deformation is given by: from which we may write: (3) The total mass m of this segment does not change, but the linear mass density changes to , given by: Page 3 of 7 Using equations (2) and (3), we may write: (4) Let us suppose now that we have a moving wheel, and we define a translational relative referential (O, x*, y*) fixed to the center O, moving to the right with velocity VO, as shown in Figure 5. In this relative referential, the point O is fixed and the ground is going to the left. The wheel rim is rotating with angular velocity (it is the same in the translational relative and absolute movements), and we see the wheel rim\u2019s point C and the tire belt\u2019s point D going to the right, with velocities VC and VD, respectively. We define the points C* and D* as fixed positions in the relative referential, i.e., these points are always in the vertical line defined by O. Let us suppose that, for the instant t = 0, the angular position of the wheel rim is zero",
            " For a complete revolution of wheel rim, we have = 2 and we obtain the time required for this revolution from equation (5): (8) In this case, to preserve the integrity of the sidewalls, one complete revolution of the wheel rim must occur in the same time of a complete revolution of the tire belt. For a complete revolution of the tire belt, the amount of mass which pass by the point D* must be equal to the tire belt\u2019s Page 4 of 7 total mass M. Using equation (7), we obtain the time required for this complete revolution: (9) As the wheel rim and the tire belt must complete one revolution at the same time, we have from equations (8) and (9): (10) As the total tire belt\u2019s mass M is equal to 0.2 R, we obtain: (11) If we change to the absolute referential shown in Figure 5, we have: (12) where V is the velocity of the non-sliding rigid disk center, as shown in Figure 1 and in equation . Constant acceleration ride case In case of vehicle\u2019s ride in constant acceleration, i.e., wheel rim under constant torque T, the wheel rim\u2019s angular acceleration is constant, as well the relative acceleration and the tire belt\u2019s deformation at point D*. Note that: - as the torque T is constant, the stress (and consequently the strain or deformation ) in the tire belt, when it starts to contact the ground, is always the same (constant); - as there is no sliding between the tire belt and the ground, the deformation will be the same along all the contact extension",
            " For a complete revolution of wheel rim, = 2 and we obtain the time required for this revolution from equation (5): (13) Also in this case, to preserve the integrity of the sidewalls, one complete revolution of the wheel rim must occur in the same time of a complete revolution of the tire belt. For a complete revolution of the tire belt, the amount of mass which pass by the point D* must be equal to the tire belt\u2019s total mass M. Using equation (7), we obtain the time required for this complete revolution: (14) As the wheel rim and the tire belt must do a complete revolution at the same time, we have from equations (13) and (14): (15) As the total tire belt\u2019s mass M is equal to 0.2 R, we obtain: (16) If we change to the absolute referential shown in Figure 5, we have: (17) where is the acceleration of the non-sliding rigid disk center, as shown in Figure 1 and in equation . If we multiply both sides of equation (17) by the time t, we obtain: (18) Note that this relationship is the same of the case of constant velocity, in equation (12). In cases when the acceleration is not constant, but it changes smoothly as in usual urban or highway trips, the equation (18) (or (12)) still may be used without significant loss of accuracy. Next, we will examine the three non-sliding \u201cslip\u201d cases: longitudinal, lateral and spin"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure16-1.png",
        "caption": "Figure 16. Normalised Stress at RSMB region (Linear analysis)",
        "texts": [
            " This is not so in the case of local non-linear analysis as contact is defined and are shown in Figure 13. Figure 14 and 15 shows comparative stress plot at RSFB region between linear and local non-linear analysis respectively. In linear analysis, high stress is observed in the fillet region of External flitch due to load transfer only through bolts and non-consideration of contact between External flitch top flange and FSM top flange. This is not so in case of local non-linear analysis as contact is defined. Figure 16 and 17 shows comparative stress plot at RSMB region between linear and local non-linear analysis respectively. Similar trend is observed in the fillet regions and high stress is shifted to other locations due to consideration of contact. In linear analysis, high stress is concentrated around bolt holes as the load transfer happens through bolts only. Figure 18 and 19 shows comparative stress plot at RSRB region between linear and local non-linear analysis respectively. Similar trend is observed in the fillet regions and high stress is shifted to other locations due to consideration of contact"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003219_i2mtc.2012.6229587-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003219_i2mtc.2012.6229587-Figure1-1.png",
        "caption": "Figure 1. Schematic diagram (not to scale) of the sensor fabrication and test device.",
        "texts": [
            "5cm x 1cm size and sonic cleaned in acetone for 20 minuets followed by rinsing again in acetone and dried at 40 \u00baC. These glass slides were This work was funded by FAIS research grant, KIT and CQU 978-1-4577-1772-7/12/$26.00 \u00a92012 IEEE then placed in a DC plasma sputtering machine and a 50nm layer of platinum(Pt) was deposited under low pressure argon(Ar) gas environment. Then the Pt deposited glass plates were masked to have a 1cm x 1cm window opened at one end and the other end kept open for terminal connection. Figure 1 shows the cross section of the device fabricated for sensor development and testing. The inside insulator strip is 0.8mm thick and two stainless steel plates of 100 m were attached to it on flat surfaces. The platinum coated glass and a platinum electrodes were placed on the stainless steal plates in a way that two platinum surfaces faces each other. In this way, now we have two platinum surfaces of 1cm x 1cm of common cross section area separated by a gap of 1mm. Both working electrodes(W1 & W2) of Hokuto Denko HZ5000 were connected to the platinum coated glass electrode of the device and counter and reference electrodes of HZ-5000 were connected to platinum plate"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003148_kem.473.875-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003148_kem.473.875-Figure1-1.png",
        "caption": "Fig. 1: a) Roboforming setup, b) basic forming strategies, c) free-form surface and d) cylinder with undercut (97\u00b0 wall angle)",
        "texts": [
            " single-point incremental forming \u2013 SPIF and two-point incremental forming \u2013 TPIF). In most cases these methods only require simple dies or supports ([4], [5]). In this regard, a robot-based incremental sheet metal forming process (roboforming) [6] is developed at the Chair of Production Systems (LPS) at the Ruhr-University of Bochum. The roboforming principle is based on flexible shaping by means of freely programmable path-synchronous movements of two industrial 6-axis robots driving universal workpiece-independent forming tools. Figure 1 a) shows the robot cell Key Engineering Materials Vol 473 (2011) pp 875-880 Online: 2011-03-28 \u00a9 (2011) Trans Tech Publications, Switzerland doi:10.4028/www.scientific.net/KEM.473.875 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: 132.174.254.159, Pennsylvania State University, University Park, USA-22/05/15,16:59:32) which consists of two KUKA KR360 robots and a clamping device",
            " The final shape is produced by the incremental inward motion of the forming tool in depth direction and its movement along the contour in lateral direction by driving either on parallel layers or on a helical path. The supporting tool holds the sheet on the backside by moving synchronously along the outer contour. The first possibility is that the supporting tool remains at the same level in thickness direction. Alternatively it may be directly opposed to the forming tool. Thereby, a predefined gap between the two hemispherical tools is achieved (Fig. 1b). Due to the kinematics of the robots and the tools\u2019 simplicity the method of roboforming is highly flexible with respect to the geometries to be manufactured by this forming process. Therefore roboforming provides a quite interesting method for industrial prototyping and low batch size production. Using the two 6-axis kinematics it is possible to form undercuts (Fig. 1d) in a multi-step strategy. This goes beyond the forming of freeform surfaces (Fig. 1c) achieved by other forming strategies. Nevertheless, up to now the resulting geometries formed in roboforming deviate several millimeters from the planned geometry. The reproducibility of these deviations can be shown by various experiments. Hence, the forming strategy with its forming parameters has a main influence on the process result. Depending on the material and the geometry being formed, parameters like the infeed, the forming velocity or the tool path especially affect the part\u2019s accuracy"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.47-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.47-1.png",
        "caption": "Fig. 2.47 4RPaPPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology R||Pa\\P\\\\Pa",
        "texts": [
            " PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34. 4PaRPPa (Fig. 2.46) Pa||R\\P\\\\Pa (Fig. 2.22g) The axes of the revolute joints connecting links 4 and 5 of the four limbs are parallel 35. 4RPaPPa (Fig. 2.47) R||Pa\\P\\\\Pa (Fig. 2.22h) Idem No. 15 36. 4PaPaPR (Fig. 2.48) Pa\\Pa\\\\P\\\\R (Fig. 2.22i) Idem No. 4 37. 4RPaPaR (Fig. 2.49) R\\Pa||Pa\\kR (Fig. 2.22j) The first and the last revolute joints of the four limbs are parallel 38. 4RPaPaR (Fig. 2.50) R\\Pa||Pa\\kR (Fig. 2.22k) Idem No. 37 39. 4RPaRPa (Fig. 2.51) R\\Pa\\kR\\Pa (Fig. 2.22l) Idem No. 15 40. 4PaRPPa (Fig. 2.52) Pa\\R\\P||Pa (Fig. 2.22m) Idem No. 34 41. 4PaPaRP (Fig. 2.53) Pa||Pa\\R\\P (Fig. 2.22n) The axes of the revolute joints connecting links 7 and 8 of the four limbs are parallel 42"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002470_amm.397-400.589-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002470_amm.397-400.589-Figure1-1.png",
        "caption": "Fig. 1 FEA model of 6205 rolling bearing",
        "texts": [
            " (ID: 141.211.4.224, University of Michigan Library, Media Union Library, Ann Arbor, USA-04/07/15,11:19:05) This paper takes a 6205 deep groove ball bearing for example. Basic sizes of the bearing are as follows: Nominal outside diameter is 52mm; Nominal inside diameter is 25mm; Raceway diameter of inner ring is 32.40mm; Raceway diameter of outer ring is 48.29mm; Average diameter is 40.38mm; Number of rolling elements is 12; diameter of rolling elements is 7.94mm; Width of the bearing is 15mm. Fig. 1 shows a FEA model of 6205 rolling bearing. According to the specific case of rolling bearing, automatic surface-surface contact pattern is chosen. Surfaces of the cage pockets, raceways of the inner and outer rings are defined as target surface. Surfaces of rolling elements are defined as contact surface. After trial calculations, the coefficients of static friction between rolling elements and inner, outer rings, cage are 0.3, 0.3 and 0.002 and dynamic friction coefficients are 0.15, 0.15 and 0",
            " ( ) ( )= cos 2 2 m bA B c e t e t m d DV V V N N N N d \u03c0 \u03b1  + = + + \u2212    (1) Provided the revolution speed of rolling elements (the same as the rotational speed of cage) is c N , then ( ) ( ) cosc b c e t e t m m V D N N N N N d d \u03b1 \u03c0   = = + + \u2212    (2) The rotational speed of rolling elements is ( )21 cos 2 m b w e t b m d D N N N D d \u03b1   = \u2212 \u2212    (3) t N is rotational speed of the inner ring; e N is rotational speed of the outer ring; b D is diameter of the rolling elements; m d is average diameter of the bearing;\u03b1 is contact angle. Termination time is determined to be 0.1s. Time interval is 100. Binary output file is ANSYS and LS-DYNA. Dynamic characteristics of the bearing are obtained based on explicit dynamics. As shown in Fig. 1, node 6146 in inner ring, node 35598 in cage, node 62812 in rolling element are selected. Fig. 2 is x-direction displacement curve of 6146, 35598 and 62812. Fig. 3 is x-direction velocity curve of 62812. The analytical solutions of speed for cage and rolling element can be calculated according to Eq. 2 and Eq. 3. Comparison between FEA results and analytical solutions is shown in Table 1: Table 1 shows that analysis results of these two methods are close and simulation results are in good agreement with analytical solutions "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001817_9781118443293.ch7-Figure7.25-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001817_9781118443293.ch7-Figure7.25-1.png",
        "caption": "Figure 7.25 Schematic image of an n-channel, polysilicon gate MOSFET, along with the symbol used for such a device in a circuit diagram.",
        "texts": [
            " The Metal Oxide Semiconductor Field Effect Transistor (MOSFET) is one of the most ubiquitous electronic components in the world with most modern computing devices containing several billion. Although most electrical engineers will have at least some familiarity with \u2212I m (Z ) \u2212I m (Z ) the theory of operation of the MOSFET it is worth reviewing the important characteristics of the device before looking at how this structure can be used to implement biological and chemical sensors. A basic knowledge of semiconductor physics, such as the difference between p-type and n-type doping, is assumed in the following discussion. The basic structure of an MOS transistor is shown in Figure 7.25 and in fact this device should probably be given a different name as the gate is fabricated from polycrystalline silicon (polysilicon) rather than metal. The substrate is p-type silicon into which are diffused or implanted two n-type regions, while the depletion regions between these are shown in red. The n-type areas are the source and drain of the transistor and are typically formed by an implantation through a mask formed by the polysilicon gate electrode in a self-aligned process. Separating the gate electrode from the silicon substrate is a very thin dielectric layer, normally of silicon dioxide",
            " In those devices the gate has now generally switched to a metal rather than a polysilicon structure, there are multiple implants to create complicated source-drain regions, and the substrate may be silicon-on-insulator or use Si-Ge to create strain in the channel to achieve improved mobilities. The easiest way to think of the MOSFET is as a switch where conduction between the source and drain contacts is dependent upon the voltage applied to the gate electrode. However, this is a four terminal device and the potential applied to the bulk or substrate also has an effect. In the n-channel device shown in Figure 7.25 the n-type source and drain regions are separated by the p-type channel region and no current will flow unless a very high voltage is applied between source and drain (VDS), causing the transistor to break down. If the voltage on the gate, which is usually referenced to source potential as VGS, is increased from 0V up to some device dependent value called the threshold voltage (VT ), then the charge on the gate will begin to repel mobile positive charge in the channel leading to the formation of a depletion region"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002275_amm.371.250-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002275_amm.371.250-Figure1-1.png",
        "caption": "Fig. 1. Industrial parts Thin walls design: a) Half-mold Lever; b) Part design \u201cLever\u201d",
        "texts": [
            " A plastic filament or metal wire is unwound from a coil and supplies material to an extrusion nozzle which can turn the flow on and off. 3D Printing is a process of making a three-dimensional solid object using an additive process, where successive layers of material are laid down in different shapes. PolyJet is a RP process that uses two or more jetting head that spray outlines of the part, layer by layer. The liquids used are photopolymers, which get cured nearly instantly by a UV lamp. In these papers is presented the study made with FDM system and \u201cDimensions sst\u201d machine only. As it may be seen from the Fig. 1 and after the \u201cCATALYST\u201d FDM system software simulations, the authors made, step by step, various design transformations concerning the industrial parts \u201dHalf-mold Lever\u201d connected with the specific capabilities of the RP technologies. There have been several attempts on simulations to determine suitable part deposition orientation for different objectives like dimensional accuracy, build time, support structure and his effect, etc. (Table 1 and Fig. 2). pressure some simulation tests using SolidWorks Simulation software were done for different materials and for different RP technologies (Table 2, Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002370_ijvas.2011.041387-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002370_ijvas.2011.041387-Figure3-1.png",
        "caption": "Figure 3 Model of the motion mechanism in the second step",
        "texts": [
            " As shown in Figure 2, this mode of crawling is performed by an articulated mechanism that one course of motion is divided to four steps, and in each step robot may be modelled by a manipulator. Moreover, it is supposed that friction between the non-moving links and ground is so much that causes these links not to slide. In this figure, the fixed links in each step are indicated with dark colour. For instance, the second step of motion course is chosen to be simulated. In this step, robot changes its configuration from state 1 to state 2, as shown in Figure 3. The angular positions of joints at beginning and end of the motion is given by Table 1. Moreover, it is desired that the joint velocities at the same times become zero. Besides, to avoid redundancy in system, the last joint in sub-mechanism is taken to be passive, i.e., \u03c44 = 0. Now, by considering identical links for robot (i.e., mk = m, lk = l, and Ik = I \u2261 ml2/3, k = 1, 2, 3, 4) and using equation (23), the equations of motion of the mechanism are obtained as follows: 2 12 13 14 12 13 141 1 22 2 12 23 24 12 23 242 2 2 13 23 34 13 23 343 3 2 14 24 34 14 24 344 4 20 15 9 3 0 15 9 3 15 14 9 3 15 0 9 3 9 9 8 3 9 9 0 36 6 3 3 3 2 3 3 3 0 C C C S S S C C C S S Sml ml C C C S S S C C C S S S \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5           \u2212     +      \u2212 \u2212         \u2212 \u2212 \u2212     1 11 2 22 3 33 4 4 7cos 1 1 0 0 5cos 0 1 1 0 3cos 0 0 1 12 cos 0 0 0 1 0 mgl N \u03d5 \u03bb\u03c4 \u03d5 \u03bb\u03c4 \u03d5 \u03bb\u03c4 \u03d5 \u03bb             \u2212           \u2212     + = +      \u2212                  (24) where (cos sin )"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002681_ijcnn.2012.6252849-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002681_ijcnn.2012.6252849-Figure1-1.png",
        "caption": "Figure 1. The concept of informationally structured space",
        "texts": [
            " copyright Robot partner can measure environment information required by the environment system, because the robot partner can move around in the living space. If our system occurs a blind spot by dynamic objects in living space, the system compensates for missing information by using the robot partner with inner sensor. Therefore, the environment surrounding people and robots should have a structured platform for gathering, storing, transforming, and providing information. Such an environment is called informationally structured space (Fig.1) [12]. The most important task in the informationally structured space is the estimation of human position, state, and behavior for natural communication. The robot partner can use the estimated information at any time. We developed a robot partner; MOBiMac shown in Fig. 2. Two CPUs are used for the interaction with human and the control of the robotic behaviors. The robot has two servomotors, eight ultrasonic sensors, a laser range finder (LRF) and a CCD camera. An ultrasonic sensor can measure the distance to objects"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002262_j.jnoncrysol.2013.03.029-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002262_j.jnoncrysol.2013.03.029-Figure2-1.png",
        "caption": "Fig. 2. Schematic presentation of ion\u2013molecule arrangement in the system.",
        "texts": [
            " The single particle distribution function is now written as f \u03b3;\u03c6\u00f0 \u00de \u00bc exp\u03b2 J \u2207\u03c8j j2\u03b7 sin2\u03b3 cos\u03c6 \u222b\u03c0 0\u222b 2\u03c0 0 exp\u03b2 J \u2207\u03c8j j2\u03b7 sin2\u03b3 cos\u03c6 sin\u03b3d\u03b3d\u03c6 : \u00f02\u00de Here, \u03b2 \u00bc 1 KBT in Eq. (2). In order to construct the free energy F, first we have to calculate the energetic contribution Hh i arising out of the ion\u2013molecule interaction in the system. If we take a snapshot of the ion\u2013molecule arrangement in the medium then we may consider their microscopical configuration to be confined in the tilt plane composed of c and the k directions as shown in Fig. 2. Here, we have tacitly assumed that the ions primarily influence the polar angle (\u03b3) of the chiral mesogen and the biaxiality of the long molecular axis (a) is taken care of in the mean field type Carr\u2013Helfrich tilting potential U \u03b3;\u03c6\u00f0 \u00de already considered above. The distance L in the figure signifies the typical length scale between the ion and the molecule up to which the molecular long axis \u201ca\u201d is influenced and our theory is not intended to describe the ionmolecular correlation beyond this length scale. The ion\u2013molecule potential energy then can be written as UIM \u03b3;\u03c6\u00f0 \u00de \u00bc \u2212q\u03bcD L2 sin\u03b3: \u00f03\u00de Here, q is the charge of the ion and \u03bcD is the molecular dipole moment as represented in Fig. 2. When Eq. (3) is averaged over the single particle distribution function f in Eq. (2), we have the energetic contribution Hh i in Eq. (1). Therefore, Fenergetic \u00bc Hh i \u00bc \u2212q\u03bcD L2 \u222b \u03c0 0 sin2\u03b3d\u03b3\u222b 2\u03c0 0 exp\u03b2 J \u2207\u03c8j j2\u03b7 sin2\u03b3 cos\u03c6 d\u03c6 \u222b \u03c0 0 sin\u03b3d\u03b3\u222b 2\u03c0 0 exp\u03b2 J \u2207\u03c8j j2\u03b7 sin2\u03b3 cos\u03c6 d\u03c6 :\u00f04\u00de Performing the azimuthal part of the integral in Eq. (4), we have Fenergetic \u00bc \u2212q\u03bcD L2 \u222b \u03c0 0 sin2\u03b3I0 \u03b3\u00f0 \u00ded\u03b3 \u222b \u03c0 0 sin\u03b3I0 \u03b3\u00f0 \u00ded\u03b3 : \u00f05\u00de Here, I0 \u03b3\u00f0 \u00de \u00bc I0 J\u03b2 \u2207\u03c8j j2\u03b7 sin2\u03b3 is the modified Bessel function of first kind. Eq"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003421_iccis.2011.126-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003421_iccis.2011.126-Figure1-1.png",
        "caption": "Fig. 1. The specifications and the effects of the various parameters on a typical UAV.",
        "texts": [
            " The rest of the paper is organized as follows: Section II presents the UAV longitudinal controller, Section III presents the experimental analysis of the system, and Section IV presents our conclusions. II UAV LONGITUDINAL CONTROLLER In this section, we developed the model for UAV longitudinal controller. This controller is used for controlling speed, pitch angle and the altitude of the UAV. The symbols used in this paper are defined in Tables 1 and 2. For the sake of illustration, we have shown these specifications in Fig. 1. 978-0-7695-4501-1/11 $26.00 \u00a9 2011 IEEE DOI 10.1109/ICCIS.2011.126 831 The controller is composed of three internal loops and two external loops. The internal loops were designed to give directive instructions for the UAV elevator and the throttle, and the external loops were designed to adjust the pitch angle. The first internal loop, based on the created error in the UAV pitch angle, directs the UAV elevator. In addition, in order to maintain the pitch angle, the first internal loop will control the elevator"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001904_9781119998914.ch10-Figure10.51-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001904_9781119998914.ch10-Figure10.51-1.png",
        "caption": "Figure 10.51 Radial dimensions of the stator",
        "texts": [
            " Thus Rshf = (Ra + Rb)/2 (10.131) where Ra = 1 2\u03c0kshf Ls + Lbs 2\u03c0kshf (Dshf / 2)2 Rb = 1 4\u03c0kshf Lb + Lbs 2\u03c0kshf (Dshf / 2)2 kshf is the thermal conductivity of the shaft, Lb is the thickness of the bearing, Dshf is the radius of the shaft, and Lbs is the distance from the bearing center to rotor mean. The Radial Conduction Thermal Resistance of Stator Teeth, Rst As both the rotor and stator core consist of layers of laminations, only the thermal conductivity in the radial direction is considered (Figure 10.51). In order to calculate the thermal resistance of the stator precisely, the stator is modeled as two parts, one as the stator yoke and the other as the stator teeth. The equivalent cylinder with a reduction factor p is used to model the stator teeth. Thus Rst = ln(rms/ris) 2\u03c0kiroLs\u03c1 (10.132) where ris the inner stator radius, rms is the inner stator yoke radius, kiro is the thermal conductivity of the stator, \u03c1 is the percentage of the teeth section with respect to the total teeth plus all slots section"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002464_ssp.198.33-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002464_ssp.198.33-Figure1-1.png",
        "caption": "Fig. 1. Details of the considered mechanism: sketch of its structure, limbs numbering, trajectory (a); constrained multibody system (b); reference tree structure of the multibody system (c)",
        "texts": [
            "159, Pennsylvania State University, University Park, USA-26/05/15,22:52:59) The force singularities can be reduced in some cases [1, 2, 13]. Moreover, the redundancy concept is a beneficial idea when difficulties are expected in the future reparations and/or services. With/after some reorganization (limited to the control mode, only), the robot can work correctly, even when some of the actuators fail. Of course, negatives are present, too. Employed incorrectly, it could effect in significant internal forces in the mechanism. It could complicate the control unit, too [2]. At present, a planar parallel mechanism is considered (Fig. 1a). Its numerical multibody model [3, 4] is tested. Rigid bodies connected by massless joints are used to model the mechanism. Limbs final points are fixed to the reference by use of constraint equations. Concerning its actuation mode, all joints are actuated in the mechanism. As user kinematics constrains are set on the platform motions, the requested joint accelerations can be estimated. To obtain the requested joint actuations, an open-loop model-based controller is postulated. Obtained accelerations are introduced in the controller dynamic model and the necessary joint torques are taken",
            "b, it leads to ( ) vuvuvPvPvAvAu FBFqMBMQBQBQ \u22c5+\u2212\u22c5\u22c5+\u2212=\u22c5\u2212\u22c5\u2212 : TA v TA u JJB \u2212 \u22c5= , (10) where: QvA \u2013 torques at the articulated/dependent joints; QvP \u2013 torques at the passive/dependent joints; BvA , BvP \u2013 columns of B in front of the articulated/dependent and passive/dependent joints. Additional simplification is possible, when the joint torques equal zero for the passive joints. As the coefficient matrix is not a square matrix, thus the right Moor-Penrose pseudo-inverse is used, and ( ) ( )vuvuA FBFAqMBMAQ \u22c5\u2212\u22c5+\u22c5\u22c5+\u2212\u22c5= ++ [ ].col vAvAAA BEAQQQ \u2212== ;) ,( : (12) where: A+=AT\u22c5 (A \u22c5 AT)-1 is the Moor-Penrose pseudo-inverse of a non-square matrix A. The algorithm leads to a particular solution where the Euclidian-norm of QA is minimal [13]. A planar 3RRR parallel manipulator (Fig. 1a), similar to the one presented in [8], is considered. In contra to classical solutions (actuations installed in limb/reference joints) all joints are actuated. Considered manipulator is modelled numerically. Joint coordinates are used as the multibody system coordinates, thus a multi-branch kinematical chain is proposed for the model. Connections of higher degrees of freedom are excluded and modelled as kinematical chains composed of massless bodies and joints. Actuators are installed in joints. A system composed of 18 bodies is proposed (Fig. 1b-c). Bodies #1 and #2 are dimensionless and massless (factious, joints linking elements, obligatory in the three degree of freedom chain between the mobile platform and the reference). The platform is a composed element, built of bodies #3\u00f76. Joints #4\u00f76 are locked at their nonzero constant values. They are followed by massless, one-dimensional bodies (with identical lengths, all). The introduced sequence facilitates platform modelling (i.e. it facilities selection of the joint positions, located at vertexes of an equilateral triangular platform, the triangle sides equal 0",
            " Bodies #7; #11 and #15 are identical (limb forearms), why bodies #8; #12 and #16 (identical, too) refer to the limb arms. Forearm and arm masses are 0.0937 kg and 0.1406 kg, respectively. Moments of inertia are 5.0765\u22c510 -5 kg\u22c5m2 and 1.6987\u22c510 -4 kg\u22c5m2 . Their nonzero sizes (x components) are 0.08 m and 0.12 m. The mass centres are in the arm and the forearm centres. Displacements at joints #1\u00f73 are taken as independent. A user trajectory is set to the coordinates (i.e. on the central point of the platform). A circle of 0.03 m is set (Fig. 1b). Platform\u2019s orientation is fixed. Constant angular acceleration (\u03b5 = 4\u22c5\u03c0 rad/s 2 ) is supposed at initial 0.5 s, and next the velocity is preserved for 0.5 s. Finally, 0.5 s period of constant deceleration (\u03b5 = - 4\u22c5\u03c0 rad/s 2 ) is supposed. Joint torques are obtain from Eq. 12. Their time evolutions are presented in Fig. 3. Discontinuities correspond to acceleration - constant velocity - deceleration switches. Obtained torques are introduced in the manipulator numerical model and time integration is performed in MATLAB [14]"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.41-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.41-1.png",
        "caption": "Fig. 2.41 4PPaRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology P\\Pa||R\\Pa",
        "texts": [
            "21u) Idem No. 11 22. 4PaRPP (Fig. 2.36b) Pa\\R\\P\\kP (Fig. 2.21v) Idem No. 5 23. 4PCPa (Fig. 2.37a) P\\C\\Pa (Fig. 2.21w) Idem No. 5 24. 4PPaC (Fig. 2.37b) P||Pa\\C (Fig. 2.21x) The last joints of the four limbs have parallel axes 25. 4CPPa (Fig. 2.38a) C\\P||Pa (Fig. 2.21y) The first joints of the four limbs have parallel axes 26. 4PaPC (Fig. 2.38b) Pa||P\\C (Fig. 2.21z) Idem No. 24 27. 4PaCP (Fig. 2.39) Pa\\C\\P (Fig. 2.21z1) Idem No. 11 28. 4PPaPaR (Fig. 2.40) P\\Pa\\\\Pa\\kR (Fig. 2.22a) Idem No. 4 29. 4PPaRPa (Fig. 2.41) P\\Pa||R\\Pa (Fig. 2.22b) The axes of the revolute joints connecting the two parallelogram loops of the four limbs are parallel (continued) Table 2.3 (continued) No. PM type Limb topology Connecting conditions 30. 4PRPaPa (Fig. 2.42) P\\R||Pa\\Pa (Fig. 2.22c) Idem No. 5 31. 4RPPaPa (Fig. 2.43) R\\P\\kPa\\\\Pa (Fig. 2.22d) Idem No. 15 32. 4PaPPaR (Fig. 2.44) Pa\\P\\\\Pa\\\\R (Fig. 2.22e) Idem No. 4 33. 4PaPRPa (Fig. 2.45) Pa\\P\\kR\\Pa (Fig. 2.22f) The axes of the revolute joints connecting links 5 and 6 of the four limbs are parallel 34"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.11-1.png",
        "caption": "Fig. 2.11 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRRP (a) and 4RPRRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R||R||R\\R\\P (a) and R\\P\\kR\\R||R (b)",
        "texts": [
            "1j) Idem No. 1 12. 4PRRRR (Fig. 2.7b) P\\R||R\\R||R (Fig. 2.1k) Idem No. 3 13. 4PRRRR (Fig. 2.8a) P||R||R||R||R\\R (Fig. 2.1l) The last revolute joints of the four limbs have parallel axes 14. 4RPRRR (Fig. 2.8b) R||P||R||R\\R (Fig. 2.1m) Idem No. 13 15. 4RRPRR (Fig. 2.9a) R||R||P||R\\R (Fig. 2.1n) Idem No. 13 16. 4RRRPR (Fig. 2.9b) R||R||R||P\\R (Fig. 2.1o) Idem No. 13 17. 4RPRRR (Fig. 2.10a) R||P||R||R\\R (Fig. 2.1p) Idem No. 13 18. 4RRPRR (Fig. 2.10b) R||R||P||R\\R (Fig. 2.1q) Idem No. 13 19. 4RRRRP (Fig. 2.11a) R||R||R\\R\\P (Fig. 2.1r) The before last revolute joints of the four limbs have parallel axes 20. 4RPRRR (Fig. 2.11b) R\\P\\kR\\R||R (Fig. 2.1s) Idem No. 3 21. 4RRRRP (Fig. 2.12a) R\\R||R||R||P (Fig. 2.1t) The first revolute joints of the four limbs have parallel axes 22. 4RRRPR (Fig. 2.12b) R\\R||R||P||R Fig. 2.1u) Idem No. 21 23. 4RRPRR (Fig. 2.13a) R\\R||P||R||R (Fig. 2.1v) Idem No. 21 24. 4RPRRR (Fig. 2.13b) R\\P||R||R||R (Fig. 2.1w) Idem No. 21 25. 4RRRPR (Fig. 2.14a) R\\R||R||P||R (Fig. 2.1x) Idem No. 21 26. 4RRPRR (Fig. 2.14b) R\\R||P||R||R||R (Fig. 2.1y) Idem No. 21 27. 4RPRRP (Fig. 2.15a) R\\P\\kR\\R\\P (Fig. 2.1z) Idem No"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002278_2013-01-1491-Figure11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002278_2013-01-1491-Figure11-1.png",
        "caption": "Figure 11. Transmission Error, Load distribution Contact stress and bending stress without micro geometry",
        "texts": [
            " A Romax model of 6 speed manual transmission has been shown in Figure 10. The Romax model of transmission has been run and analyze for the 2nd speed at defined load case. The transmission error, mesh stiffness and the contact pattern with the designed gear macro geometry has been analyzed. The micro geometry parameters like lead crowning and tip relief and, radial gap and run out (calculated) are not considered for this analysis. The load distribution during meshing of the gear tooth is not uniform and edge loading is observed as shown in Figure 11. Also, the transmission error or misalignment of the gear tooth i.e. displacement in the line of action is 2.84 micron, which reduces the efficiency and increases the noise while meshing. The load per unit length is 436.89 N/mm. The contact stress on the edge of the tooth is 1301.4 MPa and the bending stress at root is 391.73 MPa is observed as shown in Figure 11. Now, to improve the efficiency of the 2nd gear pair, micro geometry of the gear pair is optimized. The lead crowning of 10 micron and tip relief of 15 micron has been given to the drive side of gear tooth. The transmission error, load distribution, contact pattern, contact stress and bending stress with designed gear macro geometry has been analyzed with given parameters of micro geometry. The radial gap and run out due to the tilting of gear is not considered for this analysis. It is observed that with defined gear micro geometry, the transmission error or misalignment is reduced to 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002490_amm.387.328-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002490_amm.387.328-Figure2-1.png",
        "caption": "Figure. 2 Finite element mesh model",
        "texts": [],
        "surrounding_texts": [
            "Experimental modal analysis is the study of complex mechanical and engineering structural vibration. Usually measuring the vibration response of the surface of the structure by mechanical structure incentives, and then analyzing the modal parameters through series of frequency response function (FRF). In the pilot test, frequency response of the system is generally obtained through the power spectral density function such as Eq.4. ( ) ( ) ( ) FX FF S H S \u03c9 \u03c9 \u03c9 = . (4) In order to evaluate the datas quality is good or bad, that the result of Eq.5 is reliable when the coherence function value is greater than 0.95. ( ) ( ) ( ) 2 2 FX FX XX FF S S S \u03c9 \u03b3 \u03c9 \u03c9 = . (5) Here, SFX is input and output cross-power spectrum; SXX is self power spectrum of the system input; SFF is self power spectrum of the system vibration response. In the theory, any row or column of the frequency response function matrix contains all the information of the system modal parameters. Therefore, the inherent characteristics of gearbox can be obtained through measuring the frequency response function matrix in one row or one column. In this paper, the hammering method of single-input multiple-output (SIMO) is used to test gearbox structure modal. Three columns of frequency response function FRF matrix are get by moving the three direction accelerometers, and respectively knocking three directions in the fixed point. Figures. 3 and Figures. 4 respectively are test equipments and scene layout. In order to improve the accuracy of the gear system modal testing, the force percussion hammer impact velocity is increased to avoiding the the measured gear system noise signal interference and improving the signal-to-noise ratio. At the same time, avoiding the batter struck could ensure the smoothness of the measured signal waveform. Test several times with the trigger sampling, each measuring point average acquisition 5. The energizing window of a response signal should be added to accelerating the attenuation of the vibration, and to avoid the leakage of the FRF. Contrast Verification and Analysis The gearbox modal results of the CAE analysis and test measurement are shown in Table 3, and the relative error between them is also given in the following Table. From Table 3, the consistency of CAE analysis and test measurement can be summed up, and the maximum deviation of natural frequency between them is blew 15%. At the same time, it can be seen that only six of total ten bands modal are excited due to the limitations of the experimental conditions."
        ]
    },
    {
        "image_filename": "designv11_100_0002601_amr.452-453.175-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002601_amr.452-453.175-Figure1-1.png",
        "caption": "Figure 1 Sketch map of abrasion tester Figure 2 Schematic diagram of wheel-slide",
        "texts": [],
        "surrounding_texts": [
            "The severe wear or failure of wheel-slide in the flaps and slats of civil aircraft will affect the normal function of flaps and slats directly. According to report in the Aviation Safety Technology Center of Civil Aviation Administration in 2005, there were 138 incidents in the control system which caused the aircraft break-off 13 times, return 44 times, unscheduled flight 15 times, sliding back 15 times. The failures of Flaps and slats account for 64.7% of the total incidents, and according to the analysis and statistics, the most of failure mode are the jamming between slide and scroll, and the main reason of jamming is that friction and wear failure. So the research on slide and scroll has very important practical significance [1]. Rolling contact fatigue is also a common problem in aviation industry, with many important components requiring high strength and high abrasion resistance, so those components always have a long life enough to avoid being replaced in huge losses .Traditional metal materials can not meet the requirement, so in many engineering components, the coating has been used to improve the friction property of mechanical materials. In recent years, with the development of surface technology, using the composite materials including the metallic materials with coatings has becoming a trend, especially using the ceramics, plastics and other polymers as a surface coating. However, due to the coating has a poor bonding strength, the failure of coating was caused by the surface cracks and expansion of the the coating / substrate boundary and flake of coating material under the influence of contact stress. Therefore, how to design the coating and choose material is the key issue to prevent failure of the coating and substrate in system design [2-4]. So in this paper, large numbers of experiments data were analyzed by ANSYS finite element analysis software to in-depth study regularities of stress and strain at the surface and interface of coatings under different loads, different coating thickness.It has great significance for the structure of coating and substrate designed, and can guide wear test of slide and scroll wheel started effectively. Analysis of wheel-slide and coating stress Wheel-slide experimental design. Usually the needle bearing is used as rotating scroll wheel in the flaps and slats of civil aircraft, and the roller frame is fixed on the aerofoil which has the relative movement to the fuselage. To take a certain airplane wing as an example, there are 10 roller installed in the wheel-slide. 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-04/06/15,16:25:38) Because the wheel-slide size is very big, and the price is expensive, It is difficult to carry on experiments directly, So simulation test must be used to simulate the normal work of wheel-slide. so it's in addition to environmental factors ,such as loading and movement pattern should be in accordance with its normal work conditions .In accordance with these requirements, the schematic of testing machine was designed as figure 1and the schematic diagram of wheel-slide as Figure 2 In the Figure 2 the vertical load were loaded on the specimen through the wheel, and the position of wheel remains unchanged, at the same time the slide is clamped on a mobile platform and make the straight reciprocating main motion with the mobile platform to simulate working load and movement. Analysis of wheel-slide coating stress. Based on the test design requirements in subchapter 2.1, the stress distribution of the coating need be determined as an important input in order to further develop the finite element model. Because this experiment is a simulation test, the real wheel-slide should be simplified without changing the stress on the coating. The data of wheel-slide and its coating are as follows: The slide is made of Titanium alloy material (TC4), nominal composition is Ti-6Al-4V [5]; During the simulation test, The test slide have two kinds of material, namely Ti-6Al-4V, surface treatment is flame spray carbonized tungsten coating; 30CrMnSiNi2A, surface treatment is electroplating hard chrome.Before carrying on the finite element analysis, using the theoretical model to calculate the stress distribution in coating contact surface. According to the Hertz contact theory, the contact stress distribution and the size of the contact radius in coating surface were calculated [6-8]. 3 2 0 )(5784.0 \u03b7RpP = \uff081\uff09 39086.0 pRa \u03b7= \uff082\uff09 ccbb EvEv )1()1( 22 \u2212+\u2212=\u03b7 \uff083\uff09 where: the roller radius R is 25mm, the stress of contact ball vertex is P , the unit is N , the elasticity coefficient is \u03b7 , the maximum compressive stress(contact stress) is 0P , the contact radius is a . the elastic modulus isE , Poisson ratio is v , the subscript b and c are contact ball and coating respectively. In the experiment, the elastic modulus of coating is 205GPa, and the Poisson ratio is 0.25. the elastic modulus of scroll wheel is 205GPa, the Poisson ratio is 0.30, In the different loads are 50N, 100N, 500N, 1000N respectively, Corresponding stresses on the area of 10 -3 m 2 are 5\u00d710 4 Pa, 1\u00d710 5 Pa, 5\u00d710 5 Pa, 1\u00d710 6 Pa. Then with 0p and a , the following equation is adopted to calculate the stress distribution of coating: 2 0 1)(      \u2212= a r prp \uff084\uff09 Where: r is the distance from the contact center. The stress distribution as the initial load will be inputted to the finite element analysis mode, and stress distribution in coating surface is calculated by finite element analysis software. ANSYS modeling and results analysis [9]. In order to analyze stress distribution in coating surface effectively, Assumed that the coating and the substrate as a whole, the same point belong the interface of coating and substrate have a same strain, the coating and substrate have different elastic modulus and Poisson's ration. 30CrMnSiNi2A was selected as slide material when modeled by ANSYS, and surface treatment is electroplating hard chrome. The flat thickness contains coating thickness 80\u00b5m. In order to save computing resources, according to the symmetry, only choose the model of wheel 1 / 4, slide model 1 / 2,finally, the model is built as shown in Figure 3 On this basis, considering the force characteristics of the wheel-slide, the grid classification is made on the whole model, which mesh different grid partition according to the size of the stress distribution of wheel-slide as shown in figure 4. Figure5:1e5 2\u00b5m cloud Figure5:1e5 4\u00b5m cloud Figure6:1e6 6\u00b5m cloud Figure6:1e6 8\u00b5m cloud Figure7:5e5 4\u00b5m cloud Figure7:5e5 6\u00b5m cloud Figure8:4\u00b5m 1/1.5 cloud Figure8: 4\u00b5m 1/1.5 cloud The thickness of the coating selected was 2\u00b5m, 3\u00b5m, 4\u00b5m, 6\u00b5m and 8\u00b5m respectively, and the applied load was 5 \u00d7 10 4 Pa, 1 \u00d7 10 5 Pa, 5 \u00d7 10 5 Pa and 1 \u00d7 10 6 Pa respectively, and the elastic modulus was 1.5 times and 1/1.5 times of the actual value.As there are many finite element analysis result, limited to the layout only shows a representative cloud. Figure 5,5 are the representative analysis cloud with a load of 1 \u00d7 10 5 Pa and the coating thickness of 2\u00b5m, 3\u00b5m, 4\u00b5m, 6\u00b5m and 8\u00b5m, Figure 6,6 are the representative analysis cloud with a load of 1 \u00d7 10 6 Pa and the coating thickness of 2\u00b5m, 4\u00b5m, 6\u00b5m, 8\u00b5m. Figure 7, 7 are the representative analysis cloud with a load of 5 \u00d7 10 5 Pa and the coating thickness of 2\u00b5m, 3\u00b5m, 4\u00b5m, 6\u00b5m, Figure 8,8 are the representative analysis cloud with a load of 5 \u00d7 10 5 Pa, coating thickness 4\u00b5m, the elastic modulus is 1.5 times and 1/1.5 times of the actual value. According to finite element analysis results, the changes of maximum equivalent stress which under the different load conditions and different coatings thickness were analyzed and compared in Table 1-4. Table 2: Equivalent stress value under load 1\u00d7106Pa Table 3: Equivalent stress value under load 5\u00d7104Pa Table4: Equivalent stress value under load 5\u00d7105Pa load 5\u00d710 5 Pa Coating thickness DMX SMN SMX Equivalent stress value 2\u00b5m 0.263E-08 23140 0.731E+08 3\u00b5m 0.271E-08 26048 0.740E+08 4\u00b5m 0.267E-08 33058 0.672E+08 4\u00b5m 0.253E-08 38465 0.502 E+08 Based on the above equivalent stress contour maps, through contrast and analysis the conclusion is achieved: (1) Under the loads were 1\u00d710 5 Pa with the different coating thickness: 2\u00b5m, 3\u00b5m, 4\u00b5m, 6\u00b5m, 8\u00b5m, and the corresponding equivalent stress contour cloud are as shown in Figure 5 and from the figure 5 we can find out that when the coating thickness is less than 3\u00b5m, the maximum equivalent stress is located in the substrate, and the injury is occurred in substrate and then spread to the interface and coating, and that when the coating thickness is greater than 3\u00b5m, the location of the maximum equivalent stress in the coating, the first injury is occurred in the interfaces and then along the interfaces spread to the coating. The result also applies to the other loads. load 1\u00d710 6 Pa Coating thickness DMX SMN SMX Equivalent stress value 2\u00b5m 0.505E-08 45823 0.112E+09 4\u00b5m 0.532E-08 66155 0.132 E+09 6\u00b5m 0.506E-08 76931 0.100E+09 8\u00b5m 0.506E-08 76574 0.100E+09 load 5\u00d710 4 Pa Coating thickness DMX SMN SMX Equivalent stress value 2\u00b5m 0.300E-09 2371 0.144E+08 3\u00b5m 0.284E-09 2594 0.100 E+08 4\u00b5m 0.267E-09 3306 0.672E+07 (2) Contrast the contours of equivalent stress\uff0cunder the same coating thickness 4\u00b5m with the different loads :1 \u00d7 10 5 Pa, 1 \u00d7 10 6 Pa, 5 \u00d7 10 4 Pa, 5 \u00d7 10 5 Pa, and the corresponding equivalent stress contour cloud are as shown in Figure 5, Figure 6, Figure 7, and Figure 8. We can find out that the maximum equivalent stresses can be seen at the same location, and the maximum equivalent stresses have a linear relationship. The result also applies to the other coating thickness expect the 2\u00b5m, because in the elastic deformation stage, the coating thickness with 2\u00b5m has the maximum equivalent stress in the substrate. (3) Under the same load 5\u00d710 5 Pa and same coating thickness 4\u00b5m with the different elastic modulus: 1, 1.5 and 1/1.5 times of the normal value, We can find out that the location of maximum equivalent stresses are basically the same, and equivalent stress value increases with the increase of the elastic modulus, but there are no liner relationship between the stress value and elastic modulus. (4) By comparing several sets of data, changing the base coating material Poisson's ratio of coating and substrate , the location and the maximum of equivalent stress values are almost the samne, that is, Poisson's ratio of the coating and substrate have little effect on the bonding strength."
        ]
    },
    {
        "image_filename": "designv11_100_0001772_978-3-642-14019-8_3-Figure3.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001772_978-3-642-14019-8_3-Figure3.3-1.png",
        "caption": "Fig. 3.3",
        "texts": [
            " The radius vectors from the axis to the individual points of the body sweep out the same angle d\u03d5 during the same time interval dt. Thus, the angular velocity \u03c9 = \u03d5\u0307 and the angular acceleration \u03c9\u0307 = \u03d5\u0308, respectively, are the same for every point. The velocity and the acceleration of an arbitrary point P at a distance r from the axis are therefore the same as for a particle in a circular motion (see (1.25) - (1.28)): vP = v\u03d5 e\u03d5, aP = ar er + a\u03d5 e\u03d5 (3.2a) where v\u03d5 = r\u03c9, ar = \u2212 r\u03c92, a\u03d5 = r\u03c9\u0307 . (3.2b) We now consider the rotation about a fixed point A (Fig. 3.3). Let the instantaneous direction of the axis of rotation be given by 3.1 Kinematics 131 the unit vector e\u03c9. Assume that the body undergoes a rotation about this axis with the angle d\u03d5 during the time interval dt. Then all the particles of the body instantaneously move in circles. The displacement drP of an arbitrary point P is given by (see Fig. 3.3) drP = (e\u03c9 \u00d7 rAP ) d\u03d5 . (3.3) Here, the vector e\u03c9\u00d7rAP is perpendicular to e\u03c9 and rAP ; its magnitude is equal to the orthogonal distance r of point P from the instantaneous axis of rotation. We now introduce the infinitesimal vector of rotation d\u03d5 and the angular velocity vector \u03c9: d\u03d5 = d\u03d5e\u03c9 and \u03c9 = d\u03d5 dt = \u03d5\u0307e\u03c9 = \u03c9 e\u03c9 . (3.4) Then the velocity vP = drP /dt of point P follows from (3.3): vP = \u03c9 \u00d7 rAP . (3.5) It should be noted that the infinitesimal rotation d\u03d5 and the angular velocity \u03c9 = d\u03d5/dt are vectors, however, a finite rotation cannot be represented by a vector"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002867_icssem.2011.6081325-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002867_icssem.2011.6081325-Figure4-1.png",
        "caption": "Figure 4. A distribution valve backwash orifice improvement diagram.",
        "texts": [
            " BASED ON THE STRUCTURE IMPROVEMENT OF DISTRIBUTION VALVE REDUCE PRESSURE FLUCTUA nONS The backwash orifice shape and size can affect the switch-on time between the backwash pole and axis nozzle, which affects the backwash pressure fluctuations rate, and long switch-on time can leads to low pressure fluctuations rate while short switch-on time can leads to high pressure fluctuations rate. By using TT-45 (3 m 2 ) type ceramic filter as an example, the related parameters of the range of inlet pressure recoiling washing is 0.02-0.01 MPa. To prolong the switch-on time between the backwash pole and axis nozzle, we change the internal structure of the distribution valve from round to waist, which can reduce the rate of flow through the backwash orifice, and reduce pressure fluctuations. Figure 3 shows the internal diagram of the distribution valve. Figure 4 is the diagram for the distribution valve backwash orifice improvement. We design the inside round backwash orifice to waist, that is, extend original circular to waist on both sides, and expand the two waist holes for original vacuum orifice and suck orifice to large waist along two sides of circumference, which changes the backwash orifice area from 490.625 mm2 to 866.12mm2, while the other sizes don't change. With this improvement, the switch on time between the backwash pole and axis nozzle increases from original 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.56-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.56-1.png",
        "caption": "Figure 12.56 A turning plate about its diagonal on a turntable.",
        "texts": [
            " x0 z0 \u03b82 \u03b81 z2x1 B0 B1 B2 l z1 y0 x2 y1 a y2 C Figure 12.54 A turning square plate about a shaft on a turning table. 16. A Turning Plate about the Centerline on a Turntable Change the orientation of the plate of Problem 15 as shown in Figure 12.55 and solve the problem again. z0 \u03b82 \u03b81 z2x1 B1 B2 l z1 x2 y1 a y2 F F Figure 12.55 A turning plate about the centerline on a turntable. 17. A Turning Plate about Its Diagonal on a Turntable Change the orientation of the plate of Problem 15 as shown in Figure 12.56 and solve the problem again. 18. A Turning Half Square Plate about Its Side on a Turntable Change the plate of Problem 15 to half a square as shown in Figure 12.57 and solve the problem again. z0 \u03b82 \u03b81 z2x1 B1 B2 l z1 x2 y1 y2 F F a C a 21. Asymmetric Torque-Free Rigid Body (a) Prove Equation (12.411). (b) Using Equations (12.408) and (12.409), reduce Equations (12.405)\u2013(12.407) to a single equation for \u03c91. 22. Orientation of B in G In Example 745 assume the body coordinate frame B is not principal"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001269_978-3-540-33461-3_5-Figure2.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001269_978-3-540-33461-3_5-Figure2.2-1.png",
        "caption": "Fig. 2.2. Inverter: (a) delay model, (b) waveforms",
        "texts": [
            " The gates in the second row perform the complementary functions: the inverter or not gate; the nand gate; the nor gate; and the equivalence gate, equiv. For some applications, the Boolean-function model of a gate is sufficient. To study general circuit behaviors, however, it is necessary to take gate delays into account. Thus, if the gate inputs change to produce a new output value, the output does not change at the same time as the inputs, but only after some delay. Therefore, a more accurate model of a gate consists of an ideal gate performing a Boolean function followed by a delay. Figure 2.2(a) shows our conceptual view of the new model of an inverter. Signal Y is a fictitious signal that would be produced by an inverter with zero delay, the rectangle represents a delay, and y is the inverter output. Figure 2.2(b) shows some inverter waveforms, assuming a fixed delay. In our binary model, signals change directly from 0 to 1 and from 1 to 0, since we cannot represent intermediate values. The first waveform shows the input x as a function of time. The ideal inverter with output Y inverts this waveform with zero delay. If the delay were ideal, its output would have the shape y\u2032(t). 2 Topics in Asynchronous Circuit Theory 15 However, if an input pulse is shorter than the delay, the pulse does not appear at the output, as in y(t)",
            " Along this path, we have the changes y4 : 1 \u2192 0 \u2192 1, and y5 : 0 \u2192 1 \u2192 0 \u2192 1. In general, graph Ga(b) may have as many as 2n nodes, and the binary method is not practical for hazard detection; hence alternate methods have been sought [6]. We describe one such method, which uses \u201ctransients.\u201d A transient is a nonempty binary word in which no two consecutive letters are the same. The set of all transients is T = {0, 1, 01, 10, 010, 101, . . .}. Transients represent signal waveforms in the obvious way. For example, waveform x(t) in Fig. 2.2(b) corresponds to the transient 01010. By contraction of a binary word s, we mean the operation of removing all letters, say a, repeated directly after the first occurrence of a, thus obtaining a word s\u0302 of alternating 0s and 1s. For example, the contraction of the sequence 01110001 of values of y5 along path \u03c0 above is 0101. For variable yi, such a contraction is the history of that variable along \u03c0 and is denoted \u03c3\u03c0 i . We use boldface symbols to denote transients. If t is a transient, \u03b1(t) and \u03c9(t) are its first and last letters, respectively"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001672_978-3-642-39047-0_7-Figure7.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001672_978-3-642-39047-0_7-Figure7.16-1.png",
        "caption": "Fig. 7.16 An inverted pendulum system",
        "texts": [
            " The reason is that since the variable for a prismatic joint is a length di, instead of an angle, as one of the local coordinates, it will often be mixed with the dynamic parameters to be deformed together in the Riemannian metrics {wij}, causing unavailability of an effective deformer to distinguish the joint variable di from the dynamic parameters. As a first example, let us look at a well-known inverted pendulum system that consists of only two links: a linearly moving cart and a rotating pole mounted on the cart, as shown in Figure 7.16. With the joint positions x1 and \u03b82 defined as two local coordinates q = (x1 \u03b82) T for its 2-dimensional combined C-manifold, we can readily find its inertial matrixW by extracting all the coefficients of q\u0307iq\u0307j terms for i, j = 1, 2 from its kinetic energy K, i.e., W = ( m1 +m2 \u2212m2lc2s2 \u2212m2lc2s2 m2l 2 c2 + I2 ) , (7.58) where m1 and m2 are the masses of the cart and the top pole as link 1 and link 2, respectively, lc2 is the length between the revolute joint axis and the mass center of the pole, as link 2, and I2 is the moment of inertia of link 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003507_2012-01-0980-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003507_2012-01-0980-Figure5-1.png",
        "caption": "Figure 5. Anti-Slip Bearing Structure.",
        "texts": [
            " Ball stud and steel pipes are fixed to the mold and plastic is injected into the mold in order to increase the strength and stiffness for ball joint part such as a Figure 4. For ball stud in order to smooth rotate and oscillate, plastic and bearings should be well fixed. However, because using the same materials of plastic and bearing, do not be combined with each other structure. So plastic and bearing slip will occur. To prevent slip for vertical and horizontal rotation, upper and lower bearings surface were applied to the \u2018 \u2019 shape. And to combine ball stud, upper and lower bearings were designed as a locking structure as shown in Figure 5. Suspension bush is like the joints for human. And the role of the suspension bushings is absorbed the forces caused ball joint. But because bushing outer pipe material is steel or aluminum, bushing and outer pipe occur to slip and each other is lack of pull-out strength. If it is covered plastics after nurring a steel pipe, steel pipes and plastics have a strong combination and pull-out strength degradation does not occur, as shown in Figure 6. If it is designed in shown a Figure 6, current forged upper arm of production is equal to pull-out strength"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002009_978-3-319-01228-5_4-Figure4.12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002009_978-3-319-01228-5_4-Figure4.12-1.png",
        "caption": "Fig. 4.12 Underactuated manipulator with passive joint",
        "texts": [
            " In the third step, the velocity q\u0307a is computed and in the fourth step the quantity \u03c8 is evaluated. In the fifth step, the ordinary differential equation (4.127) of the internal dynamics in state space can be established and passed to the solver. With this exact inverse model, an exact reproduction of an end-effector trajectory is possible. For demonstrative purposes, the previously presented analysis and control techniques for underactuated multibody systems are applied to the control of a manipulator with a passive joint, which is shown in Fig. 4.12. The manipulator moves along the horizontal plane and consists of a cart on which an arm consisting of three homogenous links is mounted. The manipulator is described by the generalized coordinate q = [x,\u03b11,\u03b12,\u03b2]T , where x is the cart position and \u03b11,\u03b12,\u03b2 the relative joint angles of the links. The manipulator is actuated by the control input u = [F, T1, T2]T . The force F acts on the cart and the torques T1 and T2 are the driving torques acting on link 1 and link 2, respectively. The third link is connected by a passive joint to the second link which is supported by a parallel spring-damper combination with spring stiffness c and damping coefficient d"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002864_sami.2013.6480983-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002864_sami.2013.6480983-Figure1-1.png",
        "caption": "Figure 1. The F1 generating surface generates the Pk\u2019 F2 point by the tangential path of motion",
        "texts": [
            " One of such applications is the Surface Constructor (SC) that uses the original contacting and gear surface generation theory of the author, named Reaching Model. The theory, the SC system and some example applications are presented in [1]. This theory applies a special \u03ba coordinate system having curved axes \u03a6, R, T where the \u03a6 axis is parallel to the motion tracks of the Pk\u2019 points of F2 gear surface to be generated. The F2 surface which is generated by F1 in the enveloping process will be calculated in the K2 Descartes coordinate system that is connected to \u03ba as can be followed in Fig. 1. The condition of the contact of F1 and F2 surfaces is fulfilled in Pk point where the motion track of the Pk\u2019 point is tangential to F1. As the path of motion of such points is the holder of the relative velocity vectors, the n\u00b7v(1,2) = 0 necessary kinematical condition of surface contact is satisfied. The n is the common surface normal in the contact point. The discussion of the contact situation, so the relative position of the \u03a6-parallel motion path to the surface F1 reveals the possible local undercut cases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001697_v10264-012-0025-0-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001697_v10264-012-0025-0-Figure2-1.png",
        "caption": "Fig. 2. Contact conditions and geometry in face milling with workpiece [4]",
        "texts": [
            " Traction ratios of milling heads Milling heads are fitted with plates from SK, which do not conform to bending and shock loading, it is therefore necessary to know the conditions of the first image tool with the workpiece. Knowledge of these conditions allows to select geometry tool given the workpiece. Location of primary contact is an area where there is a first contact of face surface of milling tool with the workpiece. The contact of the front tool and workpiece may be in point, line or area. In addition to the type of contact, the relative position of the tool and workpiece is important Fig. 2. 4. The analyzis of the place of the primary contact of the milling head The first point of contact of a milling tool is a very important issue, because it significantly affects the durability of the cutting tool and milling quality in terms of the obtained surface roughness obtained. Roughness is generally influenced by the shape of the tip and the minor cutting edge. Inappropriate setting, as mentioned above, can be regarded as occupying an area (\"STUV\"), because already at the first contact creates a full cross section of chips and cutting wedge is subjected to considerable shock Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003336_20110828-6-it-1002.03057-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003336_20110828-6-it-1002.03057-Figure8-1.png",
        "caption": "Fig. 8. Simplified model of a double clutch.",
        "texts": [
            " In Fig. 7 the block-scheme representation of the simulative model for the system composed by n inertias is shown. Note, in particular, that by means of the power port defined by the pair (F ,\u2126), it is possible to connect the system to the models of the other mechanical elements composing the plant under investigation, as illustrated in the example considered in the following section. The working principle of the clutch used in dual clutch transmission (DCT) configurations is shown schematically in Fig. 8. It consists of two clutches that are arranged concentrically and whose friction plates are linked to the same shaft. Therefore, this system can be modeled as the three mass system of Sec. 3.2, where \u03c92, F2 are the velocity and the torque of input shaft connected to the engine, while \u03c91, F1 and \u03c93, F3 are the velocities and the torques of the shafts linked to the gearbox. The state of the clutches (engaged, slipping, or open) and therefore the power transmission from the engine to the gearbox can be separately changed by modulating the pressure between the two plates of each clutch"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.3-1.png",
        "caption": "Fig. 2.3 4RRRRR-type fully-parallel PMs with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R||R||R\\R||R",
        "texts": [
            "1 Limb topology and connecting conditions of the fully-parallel solutions with no idle mobilities presented in Figs. 2.2, 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, No. PM type Limb topology Connecting conditions 1. 4RRRRR (Fig. 2.2a) R\\R||R||R\\k R (Fig. 2.1a) The first and the last revolute joints of the four limbs have parallel axes 2. 4RRRRR (Fig. 2.2b) R||R\\R||R\\R (Fig. 2.1b) The first, the second and the last revolute joints of the four limbs have parallel axes 3. 4RRRRR (Fig. 2.3a) R||R||R\\R||R (Fig. 2.1c) The two last revolute joints of the four limbs have parallel axes 4. 4RRRRR (Fig. 2.3b) R||R||R\\R||R (Fig. 2.1c) The three first revolute joints of the four limbs have parallel axes 5. 4RRRRR (Fig. 2.4a) R||R\\R||R||R (Fig. 2.1d) The two first revolute joints of the four limbs have parallel axes. 6. 4PRRRR (Fig. 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7. 4RRRPR (Fig. 2.5a) R||R\\R\\P\\kR (Fig. 2.1f) Idem No. 5 8. 4RRPRR (Fig. 2.5b) R||R||P\\R||R (Fig. 2.1g) Idem No. 5 9. 4PRRRR (Fig. 2.6a) P\\R\\R||R\\R (Fig. 2.1h) The second and the last joints of the four limbs have parallel axes 10"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001650_978-1-4419-8113-4_16-Figure16.5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001650_978-1-4419-8113-4_16-Figure16.5-1.png",
        "caption": "Fig. 16.5 If a ball is incident with topspin and bounces with topspin then a point at the bottom of the ball has a lower horizontal speed than the middle of the ball. The ratio ex D s2=s1 is called the tangential coefficient of restitution",
        "texts": [
            " Anything that can be done to make a ball bounce faster or spin faster is at the heart of modern sports technology and the hype that surrounds it. This is especially true in golf and tennis, although similar claims are often made in relation to aluminum bats. The only way to counter the hype is to take careful measurements of ball speed and spin to determine whether there is any substance to the manufacturer\u2019s claims. Some progress has been made in this direction but a lot more still needs to be done. Suppose that a ball is incident obliquely on a horizontal surface, at speed v1, and bounces at speed v2, as shown in Fig. 16.5. The horizontal components of the ball speed before and after the bounce are vx1 and vx2, respectively. The latter speeds refer to the speed of the ball center of mass (CM). Suppose also that the ball is incident with topspin and bounces with topspin, as shown in Fig. 16.3, with angular speeds !1 and !2, respectively. A point at the bottom of the ball will have a lower horizontal speed than the CM since the bottom of the ball is rotating backward. The horizontal speeds at the bottom of the ball are s1 D vx1 R"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002299_esda2012-82282-Figure12-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002299_esda2012-82282-Figure12-1.png",
        "caption": "Figure 12 \u2212 Suspension modification: a ) original configurati on; b) minimum deflection; c) maximum deflection.",
        "texts": [
            " It should also be noted that the sum between compression and extension branches is greater for inclined suspension, due to the increase of the friction force. 4 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/75795/ on 04/09/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use To investigate the vertical load effect, independently from lateral stiffness change, the suspension springs were replaced by rigid spacers, blocking the suspension in a given deflection length. Figure 12 shows the original suspension (a), together with the modified one in the minimum (b) and maximum (c) deflection configurations. The spacers are fitted outside the suspension in the minimum extension configuration (Fig. 12c) so that the moment of inertia of the castor, with respect to the steering axis, does not change with the fork configuration; its value was identified as 0.37kgm2 by means of a free oscillation test. Castor lateral stiffness at wheel spindle level The fork length difference, between the two extreme configurations, is equal to 150 mm and implies a significant lateral stiffness difference, that was evaluated by means of static tests. The castor, constrained to the rig frame with locked steering axis, was loaded in correspondence of the wheel axis (Fig",
            " This vibrating mode is always stable if the lateral stiffness is large enough (as it happens in the airplane landing gears); however, in this case another vibrating mode may become unstable over a certain speed threshold value. This mode is characterized by higher frequencies (20\u00f730Hz), and becomes unstable at a speed values unreachable with the current test rig. Figure 29 shows a comparison between two stability curves obtained with TIRE-1 and for the two extreme deflections of the fork suspension (Fig. 12); the stiffer castor needs less damping to achieve the stability, in all the speed range. It must be noted that, by shortening the fork suspension, the castor lateral stiffness at wheel spindle level increases from 160N/mm to 270N/mm (+69%) (Fig. 15, 16); taking into account even the stiffness of the joint connecting the castor to the rig (in series with the fork one) the equivalent stiffness constant increases from 83 to 106 N/mm (+28%) (Fig.17). 9 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003081_gt2012-69967-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003081_gt2012-69967-Figure7-1.png",
        "caption": "Figure 7. Finite element mesh of motor core",
        "texts": [
            " This value was taken from a particular webbed motor for a gas compression application. Although this method requires a solid model of a webbed shaft, parameter-based modeling can be used to quickly develop various configurations from a single template model. Geometric elements selected as parameters that can be modified would consist of the base shaft diameter, web thickness and height, and the number of webs. In addition to the torsional stiffness calculation, a solid model also allows for the verification of the motor core weight and the inertia. Fig. 7 provides the meshed FEA model used to determine the rotational deflection used in the stiffness calculation. Figure 8 shows the torsional deflection of the shaft when it is subject to the face constraint at one end and the torque at the other. Figure 9 shows a close-up of the face undergoing maximum deformation. Four points are selected along the periphery of the diameter to obtain an average rotational deformation. The angular deformation is then determined from Eq. (14). l r   (14) It was observed that the torsional stiffness is generally higher as the number of spider bars as well as individual spider bar thickness increases"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003231_amm.71-78.4147-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003231_amm.71-78.4147-Figure3-1.png",
        "caption": "Figure 3 (b) gives the preload finite element model. The length, width and height of rubber are all 30 mm. Analysis requirements predicted deformation displacement state under 12 mm compression. The authors use C3D10MH unit grid and super-elastic Yeoh constitutive model to simulate deformation characteristics of rubber materials. The material parameters of Yeoh (n=3) constitutive model are: 0.145005, 0.010045, 1.48e-06, 0, 0.",
        "texts": [
            " The analysis needed to be remeshing as in [4] and its basic process is as shown in Fig.2. In especial, the secondary or even several times remeshing are needed if a remeshing scheme is not able to complete the analysis to obtain a more realistic calculation. This article describes three-dimensional remeshing the whole process of technology through the combination of ABAQUS tools and HYPERMES function. Analysis sample description. Example is a three-dimensional rubber block, and glue with upward and downward steel (Fig.3). Rubber block model\u2019s core load is the vertical displacement and grid computing is divergence when the displacement increases, to a certain extent. Therefore, the load is divided into two steps and remeshing after the first loading. Deformation analysis under no remeshing. In the simulation process, the product requires the vertical displacement of 12mm, but the Analysis cannot converge when the load is near 9mm due to the serious distortion of rubber units. The Fig.4 is given the deformation of rubber blocks under metal plates is hidden"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001110_ijtc2006-12101-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001110_ijtc2006-12101-Figure1-1.png",
        "caption": "Fig. 1. HOLE \u2013 ENTRY JOURNAL BEARING SYSTEM.",
        "texts": [
            " m-3 \u03c3 = RMS value of combined roughness, 22 bJ \u03c3\u03c3 + , \u00b5 m d\u03c8 = Coefficient of discharge for orifice \u03c6 = Attitude angle, rad yx \u03c6\u03c6 , = Pressure flow factors s\u03c6 = Shear flow factor s\u03a6 = Shear flow factor related to single surface \u2126 = Speed parameter, ( )sJrJ pcR 22\u00b5\u03c9 \u03c4 = Shear stress, ( )sJ cpR\u03c4 2 oaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?ur )(xerf = Error function, \u222b \u2212 x dyy 0 2 )exp(2 \u03c0 Subscripts and Superscripts b = Bearing J = Journal = Non-dimensional parameter The applications of hole-entry hybrid journal bearings (Fig.1) are quite wide and varied due to their superior performance such as improved load carrying capacity at zero and high speed with low energy consumption and relative simplicity in manufacturing as compared to conventional recessed or pocketed bearings [1]. Over the past several years, the incorporation of many physical effects into the analysis of fluid-film bearings has provided much more realistic performance data. In particular, the familiar assumptions of a smooth surface, isothermal operating condition and Newtonian behavior of the lubricant can no longer be employed to accurately predict the performance of fluid-film bearing system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003685_s10692-013-9486-0-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003685_s10692-013-9486-0-Figure1-1.png",
        "caption": "Fig. 1. Fig. 2.",
        "texts": [
            " The principal normal \u03bd of the helix intersects the filament axis at a right angle and coincides with its radius. Due to the properties of the helix, at any point the normal \u03bd is simultaneously a normal to the surface of the second filament - which is bent about the helical line. The series of contact points forms the axis of the twisted filament. Let the twisted filament be dissected by a plane that is perpendicular to its axis [2]. The cross section will contain two contiguous ellipses which are the cross sections of the filaments and are located at the angle \u03b1 to the twisted filament\u2019s axis (Fig. 1). The system of self-balanced internal forces acting in the cross section of the twisted filament reduces to the moments M 1 and M 2 and the forces T 1 and Q 1 . The part of the cross section that is shared by both filaments is not perpendicular to the helical axial lines of each filament. As a result, the moments M 1 and M 2 in the cross section of the twisted filament cannot be turning moments or bending moments because the forces T 1 and Q 1 are not tensile forces or shear forces, respectively",
            " (1)-(3), we write the expressions for the forces and moments acting in the filament cross section: T = q0R + T1cos\u03b1, Q = q0Rctg\u03b1 \u2013 T1sin, M\u00e8 = q0R 2(ctg2\u03b1 \u2013 1) \u2013 T1Rcos, M\u00ea =2q0R 2ctg\u03b1 \u2013 T1Rsin\u03b1. (6) In these expressions, the contact load appears as follows when it is referred to the axial line of the filaments q 0 = qcos. (7) Having designated the stiffness of one filament in torsion as B and the stiffness of the other filament in bending as H, we write relations expressing the fact that the components of the curvature and torsion are proportional to the components Fig. 1. The equilibrium of the twisted filament. Fig. 2. Internal mechanical factors. of the principal moment of the internal forces: the turning moment Mt = B\u03ba1; the bending moment Mb = H\u03ba3. Using the well-known expressions for the torsion \u03ba 1 = sin\u03b1cos\u03b1 and curvature \u03ba 3 = sin2\u03b1/R of a helix along with the obvious equality T 1 = P k /2 (where P k is the strength of the twisted filament), we obtain the formula for the contact load: . 2 sin cos sin 4 3 22 0 R B R P q k \u03b1+ \u03b1 \u03b1= (8) This parameter can also be expressed through the strength of one filament (component), which is equal to the tension at the moment of rupture T: ( ) ( ) "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003985_12.977645-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003985_12.977645-Figure5-1.png",
        "caption": "Figure 5: Components of TMP manipulation stage",
        "texts": [
            " There are three components in the precision robotic assembly system for Hohlraum assembly, namely TMP microoperation stage, Hohlraum micro-operation stage and detection system online. The composition and layout of the system are shown in Figure 3 and 4. Proc. of SPIE Vol. 8418 841819-2 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 04/21/2013 Terms of Use: http://spiedl.org/terms TMP manipulation stage is used for holding, moving and adjusting TMP. The stage is comprised of XYZ liner stages and TMP holder (Figure 5). The 3-axis liner stage with a repeatable positional accuracy of 0.5 \u03bcm can control the TMP holder motion in the XYZ direction. TMP holder is consisted by vacuum chuck, magnetic basement, 3 axis inclined stage and multi-axis force/torque sensor. The method of fixing the vacuum chuck and magnetic basement is based on magnetism and 3 axis \u201cV\u201d fix groove (Figure 6). This method can meet precise repeatable position requirement and fix quickly. There are TMP supporting ring and TMP locating pins on the work plane of the vacuum chuck"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003044_icinfa.2012.6246923-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003044_icinfa.2012.6246923-Figure4-1.png",
        "caption": "Fig. 4 Connection with guiding tracks having a locking edge",
        "texts": [
            " The handle will be released by a simple press on 5. While designing this kind of structure, the press forth and the length of the level would be highly considered to make sure the reliability of the connection. An invention represents this kind of quick connection is about a medical device connector [8]. It relates to a connection site for a medical device having a neck element with at least one guiding track. The guiding track has a lock edge for cooperative engagement with a lock protrusion of a second medical device. Fig.4 shows a medical device 1 in the form of a medical device connector 1 for connecting two medical devices. The connection site 4 comprises a neck element 5 having three guiding tracks 6 for receiving lock protrusions 3 piercing device 2. As is shown in Fig.5, the guiding track 6 exhibits Lform, comprising a first vertical section, the guiding track 6 comprises the locking edge, or barrier section, which the lock protrusions 3 of the piercing device 2 are intended to cooperate with during assembly"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001504_1-4020-3169-6_48-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001504_1-4020-3169-6_48-Figure7-1.png",
        "caption": "Fig. 7. Numerically calculated (on the top) and measured (on the botom) vibration response of the motor resulting from the 20th, 40th, 60th and 80th harmonics",
        "texts": [
            " To make the numerical and experimental data comparable, the numerical calculation of the magnetic noise and vibrations was related to the experimental investigation. Next, the exciting magnetic forces were calculated (interpolated) for the corresponding experimental loading conditions presented in Tab. 1. The interpolated exciting magnetic forces were then transferred to the structural model as the resulting magnetic forces and moments acting on the rotor and the stator, see Fig. 6. The analysis includes the first four or five harmonics of the magnetic excitation forces, depending on the rotational speed of the motor. Fig. 7 shows the mechanically deformed motor that is a consequence of the 20th, 40th, 60th and 80th harmonics of the magnetic forces. The results were obtained both numerically, using the FEM, and experimentally, using operation deflection analysis (ODA) \u2013 and good agreement was found. The deformed shape at a particular harmonic is almost identical for both motors, for the skewed and the non-skewed rotor, only the amplitude of the deformation is different. The important influence on the deformed shape of the motor is the excitation frequency of the magnetic forces, which is directly related to the rotational speed and indirectly related to the loading condition of the motor"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002529_mechatron.2011.5961097-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002529_mechatron.2011.5961097-Figure1-1.png",
        "caption": "Fig. 1. A simple two phalanxes scheme of gripping the sphere",
        "texts": [
            " This paper presents the investigation results of fixation and transportation of objects by a three finger-action gripping device (GD) from the point of view of the efficiency and reliability of these processes. The efficiency implies here the satisfaction of necessary technical requirements at the least economical, energetic and power costs. The efficiency of the workpiece gripping depends on many parameters: the material, requirements to the surface quality of the gripped element, surface and material quality of gripping jaws, number of phalanxes of gripping fingers and some other. II. INVESTIGATION OF GRIPPING THE SPHERE Let us consider the scheme of gripping the sphere, see Fig. 1. The condition of successful gripping the sphere: The workpiece will be successfully gripped (subject to the low friction of surfaces), if the point A does not exceed its outer edge. Let us write this condition: \u0430 Hence: \u0430 It is necessary to consider here the possibility of contact of the second phalanx of the finger with the surface of the gripped element: Let us consider the influence of the multiple-link character of GD on the quality of gripping. The following versions of analysis according to variable parameters can be assigned here: the number of links; the length of links; and the combination of lengths of links of gripping fingers",
            " Let us consider the loads in the moment of gripping the object: \u041c \u2013 mass of the weight; \u041c\u043a \u2013 limiting torque in the joint, when the drive of the phalanx stops increasing the load; R \u2013 radius of the sphere; \u0430\u0441 \u2013 distance between centers of the sphere and gripper. Here, the limiting factors will be: P\u043c\u0430\u0445 \u2013 maximum pressure on the surface of the object; M\u043a_\u043c\u0430\u0445 \u2013 maximum torque in the joint of the finger. The load, which promotes operations of transportation and positioning, will be considered as the useful load. The scheme with a two-phalanxes gripper (Fig.1) possesses both advantages and drawbacks (we considered them above). Vectors of forces \u04201 and \u04202 are directed only along the axis X (the hold-down action is provided), and they are oppositely directed along the axis Y. Only force P2 provides the useful load along the axis Y. The force P1 in projection on the axis Y must be minimum, but it should be present, since it promotes the fixation of the object along the axis. Now, let us determine the minimum forces to perform transportation and reliable fixation"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003849_20130911-3-br-3021.00114-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003849_20130911-3-br-3021.00114-Figure1-1.png",
        "caption": "Fig. 1. 3D model of the simulated hyper-redundant robot and details of joints and links",
        "texts": [
            " The simulation results show a smooth trajectory of the end-effector, while avoiding obstacles. The decentralized control method seems to be versatile and robust, using the redundancy and the flexibility of the manipulator. In this paper implemented algorithms of position control and anti-collision for simulated hyper-redundant robot are presented. 978-3-902823-50-2/2013 \u00a9 IFAC 572 10.3182/20130911-3-BR-3021.00114 The simulated robot has 16 joints, assembled in 8 modules of 2 joints, orthogonal to each other (fig. 1). Each joint can rotate between -90 and +90 degrees. The parameters \u03b8 (angle), d (eccentricity), a (length) and \u03b1 (twist) were obtained (table 1) according to the DenavitHartenberg model (Craig, 1990). In this work the inverse kinematics problem is addressed in two different approaches:  The Pseudo-Inverse Jacobian Method, for anti-collision;  A Heuristic Method, for the follow the leader algorithm. This method makes use of the inverse of the Jacobian matrix (1) represents the joints variation, the displacement and the inverse Jacobian matrix"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003081_gt2012-69967-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003081_gt2012-69967-Figure4-1.png",
        "caption": "Figure 4. API cross-section parameters [7].",
        "texts": [
            " API 684 METHOD API 684, first edition [2] and second edition (latest) [7], provides a method to calculate the ratio of the webbed shaft stiffness to the base shaft stiffness for a six-webbed shaft. This ratio that defines the increase in torsional stiffness is defined by Eq. 4-10 in the standard, and is reproduced below in Eq. (12). The variable definitions provided by API are shown below in Table 2. In this section, the nomenclature provided by API is followed and is different from the parameters used in the other sections of this paper. 2 223 4 24 1 16 1 1 3 1 ba b b b TL DK K DTL T D L L                            (12) The cross-sectional diagram provided as Figure 4-21 in API 684 is reproduced below in Fig. 4. Note that the variables in the previous equation do not agree with the figure below. However, one may assume that the web thickness that is labeled T in Eq. (12) corresponds to the parameter b in Fig. 4. Similarly, the base shaft diameter is provided as Db in the equation and labeled h in the figure. While the interpretations for the web thickness and base shaft diameter seem to be straight-forward, understanding the parameter L is not quite so. In Eq. (12), L is provided as the radial length of the web above the base shaft while L shows to be the overall diameter of the webbed cross-section in Fig. 4. This discrepancy in the parameter L allows for two interpretations: radial web height or overall height. The interpretation of the cross-section parameters is shown in Fig. 5 where the variables from Eq. (12) are labeled in red. For the stiffness increase calculations performed with the API equation, the interpretation from Table 2 is investigated. This interpretation yields stiffness ratios, Ka/Kb, less than one, suggesting the webs produce a decrease in torsional stiffness for the geometries considered here"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001750_9781118392393.ch4-Figure4.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001750_9781118392393.ch4-Figure4.16-1.png",
        "caption": "Figure 4.16 Helical gear with slant disc",
        "texts": [
            " The thickness of the rim is usually three or four modules, and in carburized and hardened teeth it may be over six modules. When a high-quality expensive material is determined for the teeth, only the rim is made from it, while the rest of the gear is made from grey or nodular cast iron. These two parts are press-fitted by the obligatory pin for centring. When the direction of the axial force acting on the helical gear does not change, it is desirable to slant the gear disc in a direction opposite to the force (Figure 4.16). A section through the assembly plane of a two-step reducer with cast housing and welded gears is presented in Figure 4.17. Gears roll and slide over each other, producing respective types of friction, and a part of transmitted power is expended to exceed it. Friction heats up the drive, intensifies the wear of the teeth, reduces efficiency and so on. Lubrication is vital for the faultless operation of gears. There are three aims of lubrication: Reduction of friction and power losses, Prevention of the imperil of gear load capacity, Heat dissipation caused by friction \u2013 cooling"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.62-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.62-1.png",
        "caption": "Fig. 2.62 4PaRPPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology Pa\\R||P\\Pa",
        "texts": [
            "22n) The axes of the revolute joints connecting links 7 and 8 of the four limbs are parallel 42. 4PaPaPR (Fig. 2.54) Pa||Pa||P\\R (Fig. 2.22o) Idem No. 4 43. 4PRPaPa (Fig. 2.55) P\\R\\Pa||Pa (Fig. 2.22p) Idem No. 5 44. 4PPaRPa (Fig. 2.56) P||Pa\\R\\Pa (Fig. 2.22q) Idem No. 33 45. 4PPaPaR (Fig. 2.57) P||Pa||Pa\\R (Fig. 2.22r) Idem No. 4 46. 4PaPPaR (Fig. 2.58) Pa||P||Pa\\R (Fig. 2.22s) Idem No. 4 47. 4RPPaPa (Fig. 2.59) R\\P||Pa||Pa (Fig. 2.22t) Idem No. 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.109-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.109-1.png",
        "caption": "Fig. 2.109 4PRPaRR-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 14, limb topology P||R\\Pa\\kR\\R",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003308_detc2011-48462-FigureA-1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003308_detc2011-48462-FigureA-1-1.png",
        "caption": "Figure A-1. For the Derivation of Clocking Angles on Compound Planet Gears",
        "texts": [
            " m = The number of planets that can be divided and is symmetric to other planets NP = number of teeth in the planet gear NR = number of teeth in the ring gear NS = number of teeth in the sun gear N = number of planet gears NAR = number of teeth in the ring gear (A mesh, which is in line with the sun gear) NBR = number of teeth in the ring gear (B mesh) NAP = number of teeth in the planet gear (A mesh, which is in line with the sun gear) NBP = number of teeth in the planet gear (B mesh) \u03b7BWD, = backward and forward instantaneous efficiency, \u03b7FWD respectively \u03b8C = output angle of the carrier per tooth hit (degree) \u03b8P = output angle of the planet gear per tooth hit (degree) \u03b8S = output angle of the sun gear per tooth hit (degree) \u03c6 = operating pressure angle (degree) Derivation of the Clocking Angle of Compound Planet Gears There are two gears on a compound planet gear. These two gears have to be clocked depending on the number of teeth on both planet gears and both ring gears. Assuming that there are N planet gears inside a compound planetary gear set and they are equally spaced. The Nth planet gear contacts at the operating pitch point with the ring gears and the 1st planet gear is arranged in the CCW direction in Figure A-1. The pitch angle between two planet gears is: N o360 =\u03b1 (A1) Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 9 Copyright \u00a9 2011 by ASME The pitch angle of contacting teeth on the ring gear is: )(360 N Nround N AR AR \u22c5= o \u03b3 (A2) The angle \u03b8 that the ith planet gear needs to be shifted is: \u03b1\u03b3\u03b8 \u2212= (A3) N i N Niround N AR AR AR oo 360)(360 \u22c5 \u2212 \u22c5 \u22c5=\u03b8 i =1 to N (A4) where \u03b8AR is the rotational angle that planet gear \u201cA\u201d needs to be shifted in the ring gears coordinate system"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003882_icdma.2012.156-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003882_icdma.2012.156-Figure2-1.png",
        "caption": "Figure 2. Teeth surface equations calculation",
        "texts": [
            " Rack face teeth profile is the standard basic gear profile which is shown in figure 1. The double circulararc gear end-face teeth profile equations were deduced by the normal plane teeth shape of the basic rack. This type of GB12759-1991 double circular-arc gear, each teeth is constituted by eight circular arc. But the teeth profile formed by the standard rack is not circular arc but a curve similar to the circular arc. Derivation of the double circular-arc gear teeth surface equation[7]: 978-0-7695-4772-5/12 $26.00 \u00a9 2012 IEEE DOI 10.1109/ICDMA.2012.156 658 As shown in figure 2, when the plane of the rack section and the gear pitch cylinder roll with each other, rack teeth envelopes gear teeth surface. To establish the equation of the teeth surface, four coordinate systems are set So---Space fixed coordinate system, Sn---Moving coordinate system fixed with the normal plane of basic rack, Ss---Moving coordinate system fixed with the basic rack end-face, S---Moving coordinate system fixed with gear. Assume the coordinate value of a point on the basic teeth profile: 0 cos sinnx nz Lny X (1) Where: X---Above yn-axis, take a positive value, otherwise a negative value, L---The distance between teeth profile center and xn-axis, in the positive direction of yn-axis, take a positive value, otherwise a negative one, \u03b1---Profile angle, counterclockwise direction is positive, \u03c1---teeth profile radius"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002485_s1068366613060123-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002485_s1068366613060123-Figure3-1.png",
        "caption": "Fig. 3. (a) Balls and (b) block with grooves.",
        "texts": [
            " In this tribometer, contact pairs, along with a heater and a temperature sensor that are located very near to the zone of con tact, can rotate at a required angle \u03b1 with regard to the horizontal plane. The angle of inclination of the rotat able plane is mechanically implemented with an accu racy of up to 1'. The measuring system is isolated from the zone of heating. These experiments were aimed at determining the effect of the temperature on the static coefficient of friction under the rolling of balls with various radii over grooves with various radii of the generatrix. Figure 3 shows photos of the balls and the block with grooves. The weight of the balls was 0.004\u20130.708 N and their radii r were 2.32\u201313.0 mm. Both the balls and the block were made of ASTM A 295/52100 bearing steel (an analog of ShKh 15 steel) with a Rockwell hardness of 62\u201366 HRC. The surface roughness of the balls was 452 JOURNAL OF FRICTION AND WEAR Vol. 34 No. 6 2013 TODOROVIC et al. Ra = 0.002 \u03bcm. The block was machined to produce different radii of the section of the groove profile (R1 = 2.5, R2 = 5"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003511_amm.198-199.171-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003511_amm.198-199.171-Figure3-1.png",
        "caption": "Fig. 3 The Overall Structure Design of the Tennis Vehicle",
        "texts": [
            " Based on that, the concept of the tennis vehicle is proposed, intending to design a new type of multi-function tennis equipment for the tennis practitioners. In order to facilitate the use, this paper designs a tennis vehicle which contains the function of serving and picking up tennis balls. In the running of the vehicle, the scattered tennis balls are collected by the picking mechanism. When the collection of balls is put into the serving mechanism, the exercisers can control the serving mechanism through the remote controller, and do tennis practice. The overall structure design of the tennis vehicle is shown in Fig. 3. drive the front wheel to rotate through the surface friction, and then it transmits rotary force to the drum of the picking mechanism which drives tennis balls to scroll up. It achieves the goal of collecting balls that picking up them into the collection box. The achievement of the tennis vehicle is shown in Fig. 4. The serving mechanism, associated control circuit and power source are all fixed in the back of the tennis vehicle, which is convenient for unified management. When the collection of balls is put into the serving mechanism, it can provide the tennis-serving practice"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure22-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure22-1.png",
        "caption": "Fig. 22 Toe/Heel chamfer limits for helical gears",
        "texts": [
            " 21) are required to orient the EM at any point along the tooth edge: Vo: N X T ,\u2212\u2212\u2192 Tool: the tool vector,\u2212\u2212\u2212\u2192 Trans: axis about which vector \u2212\u2212\u2192 Tool is rotated by the Pivot Angle,\u2212\u2212\u2212\u2192 Pivot : axis about which vector \u2212\u2212\u2192 Tool is rotated by the local helix angle. \u2212\u2192 Vo is obtained from the cross product of N and T . Again, vector \u2212\u2212\u2212\u2192 Trans is obtained by pivoting T about \u2212\u2192 Vo by \u03c0 2\u2014the local pressure angle; vector \u2212\u2212\u2212\u2192 Pivot is obtained from the cross product of \u2212\u2212\u2212\u2192 Trans and \u2212\u2192 Vo; then, pivoting \u2212\u2212\u2212\u2192 Pivot about \u2212\u2212\u2212\u2192 Trans by the user inputted Pivot Angle gives vector \u2212\u2212\u2192 Tool which is then pivoted about axis \u2212\u2212\u2212\u2192 Pivot to account for the local helix/spiral angle. We also see in Fig. 22 that one tooth end typically cannot be reached easily because of the risk of interference between the tool spindle and the turn table. Then, either the Pivot Angle must be reduced or an alternate solution is required, which is presented further in the text. Spiral bevel gears have the most complex shapes in terms of tooth flank topography and blank. These gears therefore call for a general solution that is applicablewhatever the local spiral angle, face width, generating process, module or pressure angle"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003187_amm.483.382-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003187_amm.483.382-Figure2-1.png",
        "caption": "Figure 2(a) the main view of fixed brush Figure 2(b) the left view of fixed brush",
        "texts": [
            " Most domestic Transmission and Distribution Engineering Company did not conduct a more thorough clean-up before oil-soaking rope. Therefore, the result of the maintenance is unsatisfactory. 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: 141.211.4.224, University of Michigan Library, Media Union Library, Ann Arbor, USA-11/07/15,07:08:45) Development and design of wire rope decontamination device The structural design of the fixed brush. The structure of fixed brush is shown in Figure 2. Seen from the figure\uff0cThree adjustable brush is mounted on a fixed semicircular plate, The brush has a threaded adjustment mechanism along the radial rope, and brush is connected with the bracket 1 through a pin and a spring. The mechanism can adjust the brush with the distance from the wire rope, it adapts to different sections of the size of wire rope, and it can adjust pressure of brush on the rope. At the same time it ensures the wire connectors passable [2] . It can be seen from Figure 2 that two long brushes are mounted on a flexible plate 2. Steel tilt angle along the rope running direction, so that the brush produces a certain pressure on the rope. Elastic plate should ensure the wire connectors passable. The structure is simple, but thickness and flexibility of spring plate have certain requirements. Therefore it is necessary to simulate and analyze this elastic plate. Simulation analysis of elastic plate. The structure of flexible steel is shown in Figure 3. Material is 60Si2Mn, the force of 10N are subjected on A and B, Assuming that the force is the brush pressure on the rope"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003659_amr.228-229.106-Figure5-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003659_amr.228-229.106-Figure5-1.png",
        "caption": "Fig. 5 Helical gear 1 forces the virtual helical Fig. 6 The virtual helical rack profile plane ABCD rack profile plane ABCD to translate right forces helical gear 2 to rotate about O2O2\u2032",
        "texts": [
            " If spiral involutes lie on certain surfaces, the surfaces certainly can be used as gear profiles with the constant transmission ratio. In fact, spiral involutes do exist on helicoids. For meshing spiral involutes, k and c should satisfy conditions as follows: p1 p1p2 11 cos sin costgtg \u03b1 \u03bb \u03bb\u03b1\u03b1 \u22c5 \u22c5\u2212 == ck (21) p2 p2p1 22 cos sin costgtg \u03b1 \u03bb \u03bb\u03b1\u03b1 \u22c5 \u22c5\u2212 == ck (22) Where \u03b1p1 or \u03b1p2 is the pressure angle on the pitch circle of the involutes. The angle between two axes is \u03bb, and the inclination angles of profile plane are respectively \u03b81 and \u03b82. Cylindrical spiral involute gears Generation of tooth profiles. As depicted in Fig.5, helical gear 1 rotating about O1O1\u2032 forces plane ABCD, a virtual common-rack-cutter surface to translate rightwards. The pressure angle in the transverse plane is \u03b1p1, and the helix angle is \u03b1p2. Simultaneously plane ABCD forces helical gear 2 to rotate about O2O2\u2032 that is perpendicular to O1O1\u2032 (See Fig.6). The pressure angle in the transverse plane for helical gear 2 is equal to \u03b1p2, and the helix angle of helical gear 2 is equal to \u03b1p1. wedge (See Fig. 8). What contribute to meshing of gear 1 and gear 2 are the engaged spiral involutes generated by BD or HF"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003772_iet-opt.2011.0058-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003772_iet-opt.2011.0058-Figure1-1.png",
        "caption": "Fig. 1 CCD transient temperature test system",
        "texts": [
            "0058 the powder is found in the handbook, the related parameters of the powder are in Table 1. An HRPS-IIIA SLS machine is used to carry out the experiment. Output maximum power of the CO2 laser is 50 W, accordingly only low melting point powder is sintered. In this paper, 95% of the total volume of sintered powder is polystyrene and the rest is Al2O3-coated ceramic powder. The size of the cuboids prototype of simulation is 200 mm \u00d7 40 mm \u00d7 6 mm. For matching the fast laser scanning speed, a testing temperature system (Fig. 1) with a high-speed video camera (model: AOS X-PRI) is proposed, its colourful area array CCD pixel is 800 \u00d7 600 and frame frequency is 1000 fps. A neutral colour filter with 50% transmissivity is used for anti-saturation. The system is calibrated by a blackbody. It follows the basic principle of testing temperature based on three wavelength colours method (i.e lR \u00bc 0.7 mm, lG \u00bc 0.546 mm and lB \u00bc 0.435 mm). The equation for solution of temperature is provided as follows [19] (see (13)). Fig. 2a shows the collected image and the divided three colours, the colour R is shown in Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002278_2013-01-1491-Figure4-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002278_2013-01-1491-Figure4-1.png",
        "caption": "Figure 4. Tolerance stackup path for maximum radial gap",
        "texts": [
            " The needle roller bearing is mounted on the outside diameter of 2nd gear bush which is press fitted with the shaft outside diameter and rotates with the speed of shaft. The nominal dimensions with maximum and minimum tolerance of all these components are given in Table 1. The first step is to calculate the maximum radial gap between 2nd gear bore diameter and the outer diameter of needle roller bearing as per the layout shown in Figure 3. The stackup path followed to find the maximum gap is shown below in Figure 4. To calculate the maximum radial gap between the gear and the needle roller bearing, subtract the minimum 2nd gear bush outer diameter and minimum roller diameter from the maximum of 2nd gear bore diameter. From the dimensions given in Table 1, the maximum radial gap between the gear and the needle roller bearing is 0.100 and the minimum is 0.050 as shown in Figure 5. The maximum tilting of the 2nd gear at maximum radial play increases the gear mesh misalignment. The sketches used to calculate the 2nd gear maximum tilting (in milli radians) is shown in Figure 6"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002047_978-1-4471-5110-4_6-Figure6.16-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002047_978-1-4471-5110-4_6-Figure6.16-1.png",
        "caption": "Fig. 6.16 Problem 6.3",
        "texts": [],
        "surrounding_texts": [
            "If curl P = \u2207\u00d7P = 0 the force P is conservative. The expression \u2207\u00d7P is calculated with MATLAB: rotP_ = curl(P_,[x y z]); and the results is: curl(P_)=[0, 0, 0] The potential energy is calculated with V (x, y, z) = \u2212 \u222b P \u00b7 dr + C = \u2212 \u222b [\u2212kxdx \u2212 kydy + (k R \u2212 kz \u2212 mg)dz] + C V (x, y, z) = k x2 2 + k y2 2 + k (R \u2212 z)2 2 + mgz + C, (6.8) where C is an arbitrary constant known as the constant of integration. The equation of the sphere can be written as x2 + y2 + z2 = R2 or z = \u221a R2 \u2212 x2 \u2212 y2, (6.9) and with Eq. (6.8) the potential energy function of x and y is calculated in MATLAB with: V1=-int(P_(1)); V2=-int(P_(2)); V3=-int(P_(3)); V=V1+V2+V3+C; fVxy=subs(V,z,sqrt(R\u02c62-x\u02c62-y\u02c62)); Vxy=simple(simplify(Vxy)); and V (x, y) is obtained as V (x, y) = k R2 + (mg \u2212 k R) \u221a R2 \u2212 x2 \u2212 y2 + C. The partial derivative of the function V (x, y) with respect to x and y are \u2202V \u2202x = \u2212 (mg \u2212 k R) x\u221a R2 \u2212 x2 \u2212 y2 , \u2202V \u2202y = \u2212 (mg \u2212 k R) y\u221a R2 \u2212 x2 \u2212 y2 . The equilibrium positions of the particle are obtained from \u2202V \u2202x = 0, \u2202V \u2202y = 0. In MATLAB the equilibrium positions are obtained with: dVxydx = simple(diff(Vxy,x)); dVxydy = simple(diff(Vxy,y)); xe=solve(dVxydx,x); ye=solve(dVxydy,y); ze=solve(xe\u02c62+ye\u02c62+z\u02c62-R\u02c62,z); The results for the equilibrium positions are M1(0, 0, R) and M2(0, 0,\u2212R). Example 6.5 A particle P of mass m is on a circle of radius R as shown in the Fig. 6.13. The circle is on a vertical plane xy. Find the equilibrium positions of the particle. Solution The independent variable is the angle \u03b8. The position of the particle P is x = R*cos(theta); y = yN+R+R*sin(theta); r_ = [x y]; where yN is the y coordinate of the lower end N of the circle. The gravity is the only force acting on the particle and the potential energy is calculated with: dr_=diff(r_,theta); G_ = [0 -m*g]; V = -int(G_*dr_.\u2019); fprintf(\u2019V=%s + C\\n\u2019, char(V)) The MATLAB expression for the potential energy is: V=R*g*m*sin(theta) + C where C is a constant of integration. The equilibrium positions are calculated from the equation: dV = diff(V,theta); thetae=solve(dV,theta); theta1=thetae; theta2=theta1+pi; The equilibrium position are the points M and N as shown in Fig. 6.13: theta1 = pi/2 and theta2 = (3 \u2217 pi)/2. The equilibrium stability is verified with the second derivative of the potential energy: d2V = diff(dV,theta); d2V1=subs(d2V,theta,theta1); d2V2=subs(d2V,theta,theta2); and the MATLAB results are: d2V/d(theta)\u02c62=-R*g*m*sin(theta) for theta1 => d2V/d(theta)\u02c62=-R*g*m for theta2 => d2V/d(theta)\u02c62=R*g*m The equilibrium position \u03b8 = 3 \u03c0/2, position N , is a stable equilibrium because d2V/d\u03b82 = Rgm is positive. 6.6 Problems B A C 1 2 k \u03b8 m, l m, l F A B A C \u03b8 m, l l k k Fig. 6.18 Problem 6.5 B A k O C F \u03b8 Fig. 6.19 Problem 6.6 B A k \u03b8 m, l k Fig. 6.21 Problem 6.8 \u03b8 l m, l m, l m, l F B A C D k 1 2 3 0 0 by the torsion spring is M = K \u03b8, where \u03b8 is the relative angle between the links at the joint. Determine the minimum value of K which will ensure the stability of the mechanism for \u03b8 = 0. 6.8 Figure 6.21 shows a four-bar mechanism with AD = l. Each of the links has the mass m (m1 = m2 = m3 = m) and the lenght l (l1 = l2 = l3 = l). At B a vertical force F acts on the mechanism and the spring stiffness is k. The motion is in the vertical plane. Find the equilibrium angle \u03b8. Use the following numerical application: l = 15 in, m = 10 lb, F = 90 lb, and k = 15 lb/in. Select an unextended (initial) length L0 for the spring. 6.9 A particle of mass m can move freely in space. The potential energy V of the particle at x = l, when the particle is subject to a vertical force F = ax2 + bx + c, is V = s. Find the equilibrium positions of the particle. For the numerical application use a = 1, b = \u22123, c = 0, l = 0 m, and s = 5 J. 6.10 A bar of mass m and length l is supported by a vertical wall and a point at O , as shown in Fig. 6.22. Find the equilibrium positions of the bar. For the numerical application use l = 0.5 m, a = 0.1 m, m = 1 kg, and g = 9.81 m/s2."
        ]
    },
    {
        "image_filename": "designv11_100_0001253_robot.2005.1570279-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001253_robot.2005.1570279-Figure1-1.png",
        "caption": "Fig. 1. Relevance of the different arm and hand joints.",
        "texts": [
            " Each of the n-1 upper levels of the tree below the root node corresponds to a certain gripper finger, and each of the m following tree levels corresponds to a certain joint of the robot, where both the different fingers and the different joints are ordered according to a previously defined hierarchy. This hierarchy should reflect the relevance of each finger and joint to grasp quality. The relevance can be established following any application specific criteria. Normally, those joints closer to the contact point (closer to the end effector) should be considered more relevant, as Fig. 1 shows. However, any other criteria would also be applicable. For understandability reasons, the following paragraphs describe the generation of the whole inverse kinematics tree; afterwards it will be explained how only a small subset of the tree needs to be computed, thus highly reducing the computational load. First, the generation of the upper subtree (the one corresponding to the first n levels, one per robot finger) will be explained. Once the hierarchy is established, this subtree is generated starting from a root node"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001897_9781118516072.ch2-Figure2.34-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001897_9781118516072.ch2-Figure2.34-1.png",
        "caption": "Figure 2.34. Open-circuit saturation characteristic of the synchronous generator used for eval-",
        "texts": [
            " Hence, only the magnetizing reactances Xmd and Xmq saturate and their saturation is determined by the air gap flux linkages. The saturated values of the mutual inductances are determined by Lmd \u00bc KSdLmd;unsat Lmq \u00bc KSqLmq;unsat (2.120) where KSd and KSq represent the degree of saturation in the d-axis and q-axis, respectively, and Lmd,unsat and Lmq,unsat are the unsaturated values of Lmd and Lmq. The factor KSd can be determined from the OCC. Let us consider an operating point marked with \u201ca\u201d on the OCC to which the field current If ;a and flux ca corresponds (Figure 2.34a). Taking into account equation (2.113) of the saturation factor and the similarity of triangles, it results KSd \u00bc ca ca \u00fe cS \u00bc If ;0 If ;a (2.121) Regarding the determination of the degree of saturation in the q-axis, the following issues are taken into account: For salient pole generators (hydrogenerators), since the closing path of the flux is mainly air, the iron saturation along the q-axis is insignificant and therefore it can be approximated that KSq \u00bc 1 for all loading conditions. For round rotor generators (turbogenerators), magnetic saturation exists also in the q-axis, and the factor KSq should be determined from the no-load saturation characteristic along the q-axis",
            " measurement) it can be approximated, with good accuracy that KSq \u00bc KSd. This approximation is equivalent with the assumption that the air gap and themagnetic path reluctance, respectively, are nonuniform along the rotor circumference. Taking into account the above-presented issues, it results that, in order to represent the effects of saturation, it is necessary to identify a suitable mathematical function, which can quantify the deviation of OCC from the air gap line. In this regard, the OCC is divided into three segments (Figure 2.34b) [1]: Segment I, corresponding to the interval in which saturation phenomenon does not occur, is characterized by values of the flux linkages lower than the threshold value cI, which is usually 0.8 p.u. In this case, it results cS \u00bc 0 (2.122) Segment II, corresponding to a partially saturation of the magnetic iron, is characterized by values of the flux linkages greater than cI and lower than cII. The threshold value cII over which full saturation of the magnetic iron occurs is usually 1.2 p.u. In this case, cS can be expressed by an exponential function: cS \u00bc AS e BS ca cI\u00f0 \u00de (2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001737_978-1-4614-8544-5_1-Figure1.1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001737_978-1-4614-8544-5_1-Figure1.1-1.png",
        "caption": "FIGURE 1.1. Cross section of a tire on a rim to show tire height and width.",
        "texts": [
            " Tire and Rim Fundamentals Tires are the only component of a vehicle to transfer forces between the road and the vehicle. Tire parameters such as dimensions, maximum loadcarrying capacity, and maximum speed index are usually indicated on its sidewall. In this chapter, we review some topics about tires, wheels, roads, vehicles, and their interactions. 1.1 Tires and Sidewall Information Pneumatic tires are the only means to transfer forces between the road and the vehicle. Tires are required to produce the forces necessary to control the vehicle, and hence, they are an important component of a vehicle. Figure 1.1 illustrates a cross section view of a tire on a rim to show the dimension parameters that are used to standard tires. The section height, tire height, or simply height, , is a number that must be added to the rim radius to make the wheel radius. The section width, or tire width, , is the widest dimension of a tire when the tire is not loaded. Tires are required to have certain information printed on the tire sidewall. Figure 1.2 illustrates a side view of a sample tire to show the important information printed on a tire sidewall"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003072_amr.744.185-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003072_amr.744.185-Figure2-1.png",
        "caption": "Fig. 2 Slider-crank Mechanism with a Revolute Joint",
        "texts": [
            " )`,,,( ),(),(2 ))()((2))()(,(2 )()()()( 10 0 0, 2 2 0000 2 2 min n n k n ki kikikkD DnnDnnD DD aaaQ aafaf xaxaxaxaff xfxxfxg \u22c5\u22c5\u22c5= +\u2212= \u22c5\u22c5\u22c5+++\u22c5\u22c5\u22c5+\u2212= \u2212=\u2212 \u2211 \u2211 = = \u03a6\u2208 \u03d5\u03d5\u03d5 \u03d5\u03d5\u03d5\u03d5 \u03d5 \u03d5 (8) ,0= \u2202 \u2202 ia Q ni ,,0 \u22c5\u22c5\u22c5= (9) \u2211 = = n k ikik fa 0 ),,(),( \u03d5\u03d5\u03d5 ni ,,0 \u22c5\u22c5\u22c5= (10) It is expressed in matrix form as:           \u22c5\u22c5\u22c5=           \u22c5\u22c5\u22c5           \u22c5\u22c5\u22c5 \u22c5\u22c5\u22c5\u22c5\u22c5\u22c5\u22c5\u22c5\u22c5 \u22c5\u22c5\u22c5 Dn D nDnnDn DnD f f a a ),( ),( ),(),( ),(),( 00 0 000 \u03d5 \u03d5 \u03d5\u03d5\u03d5\u03d5 \u03d5\u03d5\u03d5\u03d5 (11) The journal surface is divided into several sectors( ,,, 210 SSS \u2026 nn SS ,, 1\u2212 ), the time is divided into several cycles ( ,,, 210 TTT \u2026 nn TT ,, 1\u2212 ), each cycle as the journal from a limit position to another position limit, thus, it is assumed that the extrapolation function h(t) : btath +=)( (12) The wear date drawn from the incremental model in a cycle as the following Table1: the normal equation of For any sector can be expressed as:             =                  \u2211 \u2211 \u2211\u2211 \u2211 = = \u2212= = p i kimi p i ki m i mi p i mi p i mi hT h b a TT Tp 0 0 0 2 0 0 (13) the extrapolation function of wear can be obtained by solving the coefficients a, b. The slider-crank drive system is shown in Fig.2, the clearance is assumed between the revolute joint connecting the crank and the link. The parameters used in the dynamic simulation are listed in the Table2. rotates 100 circles as a data collection point, the wear date of sectors of 2S , 6S , 10S at the 5 collecting point as the Table3. The extrapolation function of the sectors 2S , 6S , 10S can be obtained according to the Eq. (3).The wear volume calculated from extrapolation function compared with the incremental model as Fig.4. It can be founded that the extrapolation function curve is very close to the wear increment curve with the computational cost greatly reduced"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.85-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.85-1.png",
        "caption": "Fig. 2.85 4PaPaRPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 42, limb topology Pa||Pa\\R||Pa",
        "texts": [],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002467_j.elstat.2011.08.002-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002467_j.elstat.2011.08.002-Figure2-1.png",
        "caption": "Fig. 2. Equipotential lines.",
        "texts": [
            " The ball jumps up and adheres to the probe due to the electrostatic interaction [9] between the electrode with a hemispherical tip and the ball. The ball adheres to the probe tip. The ball is released from the upper electrodewhen the gravitational force exceeds the holding force. This catch and release mechanism can be controlled by controlling the applied voltage. The analysis is performed for the two cases described above with the ball positioned on the center axis of the electrode with a hemispherical tip. The probe is initially in electrical contact with the lower electrode. Fig. 2 shows the electrostatic potential around the ball. 3.2. Dielectric film below electrode with hemispherical tip Electric field analysis was also performed for the case when a dielectric film is placed beneath the hemisphere electrode (see Fig. 1(b)), so that the ball can be attached below the probe tip. The dielectric film maintains the induced charge on the ball so that the ball can continue to be captured on the dielectric film. Fig. 3 shows the equipotential lines around the ball for the case when there is a dielectric film below the electrode with a hemispherical tip"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001110_ijtc2006-12101-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001110_ijtc2006-12101-Figure2-1.png",
        "caption": "Fig. 2. BEARING GEOMETRY AND SURFACE PROFILE.",
        "texts": [
            " Assuming both journal and bearing surfaces having the same surface pattern (i.e. bJ \u03b3\u03b3 = ), the shear flow factor ( s\u03c6 ) is expressed as [4] srjs V \u03a6\u2212= )12(\u03c6 (2b) where s\u03a6 is expressed as [4] 2)(3)(21 1 )( hh s ehA \u039b+\u039b\u2212\u039b=\u03a6 \u03b1\u03b1\u03b1 for 5\u2264\u039bh )(25.0 2 h s eA \u039b\u2212=\u03a6 for 5h >\u039b (2c) Copyright \u00a9 ASME 2006 l=/data/conferences/ijtc2006/71417/ on 07/20/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Dow 32121 and ,,, \u03b1\u03b1\u03b1AA are constants and can be obtained from Patir and Cheng [4]. For the journal bearing system shown in Fig.2, assuming Gaussian distribution of surface heights, the expression for average fluid-film thickness ( Th ) in fully lubricated (i.e. for 3\u2265\u039bh ) and partially lubricated (i.e. for 3<h\u039b ) regions is expressed as [13]      <\u039b \u039b +              \u039b + \u2265\u039b = \u039b\u2212 3hfor 2 1 2 1 2 3hfor 22)( hT eherfh h h \u03c0 (3) where h is the nominal fluid-film thickness i.e. the fluid-film thickness of a smooth journal bearing system and is expressed as \u03b1\u03b1 sincos1 JJ ZXh \u2212\u2212= (4a) where \u03c6\u03b5 sin=JX and \u03c6\u03b5 cos\u2212=JZ (4b) For an orifice restrictor, the equation of flow in non-dimensional form is expressed [14] as ( ) 21 2 1 cSR pCQ \u2212= (5) For the thermal analysis, the flow of lubricant between two rough surfaces can be modeled by an equivalent flow model as shown in Fig.2c [12]. The equivalent model is defined as two smooth surfaces separated by a clearance equal to the average gap ( Th ). Based on the equivalence of flows through the average fluid-film thickness and through the local fluid-film thickness, a group of new pressure flow factors ( ) and shear flow factor (yx \u03c6\u03c6 \u2032\u2032 , s\u03c6\u2032 ) can be derived [12] as x T x h h \u03c6\u03c6 3 3 =\u2032 ; y T y h h \u03c6\u03c6 3 3 =\u2032 and ss \u03c6\u03c6 =\u2032 (6) The mean or expected velocity components can be obtained by modifying the Poiseuille and Couette terms in the expression of local velocity components using the new flow factors and are expressed in non-dimensional form as \u222b\u222b\u222b\u222b \u039b \u2032\u2126 + \u2126 +         \u2212 \u2202 \u2202\u2032= z T s zzz xT zd Fh zd F zd F F zdzphu 000000 1 0 2 \u00b5 \u03c6 \u00b5\u00b5\u00b5\u03b1 \u03c6 (7a)         \u2212 \u2202 \u2202\u2032= \u222b\u222b zz yT zd F F zdzphv 00 1 0 2 \u00b5\u00b5\u03b2 \u03c6 (7b) The fluid-film velocity component across the fluid-film is obtained from the continuity equation and is expressed in nondimensional form as 4 nloaded From: http://proceedings"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001702_978-3-642-38256-7_3-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001702_978-3-642-38256-7_3-Figure1-1.png",
        "caption": "Fig. 1. A screen-printed three-electrode system. (1) working electrode; (2) reference electrode; (3) counter-electrode.",
        "texts": [
            " ccNiR was purified from Desulfovibrio desulfuricans ATCC 27774 cells grown in nitrate, as previously described by Almeida and co-workers [9]. For the optimization studies, a conventional three-electrode electrochemical cell was used, with an Ag/AgCl reference electrode, a Pt counter electrode (both from Radiometer) and a home-made working electrode made of pyrolytic graphite disks (4 mm diameter) and modified with the enzyme/ink layer. The characterization of the optimized electrode was performed after replacing the previous system by carbon paste screen-printed electrodes (CPSPEs) with a three electrode configuration (Fig. 1), including an Ag/AgCl pseudo-reference, a graphite paste counter electrode and a graphite paste working electrode (3.1 mm diameter). The CPSPEs were fabricated at CIDETEC facilities, as described by Ochoteco and co-workers [17]. The one-compartment electrochemical cell containing 0.1 M KCl in 0.05 M Tris-HCl buffer, pH 7.6 as supporting electrolyte, was thoroughly purged with Argon before each experiment. Measurements were performed with a potentiostat Autolab PSTAT 12 (Eco-Chemie) monitored by the control and data acquisition software GPES 4"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003778_978-3-642-17234-2_4-Figure4.11-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003778_978-3-642-17234-2_4-Figure4.11-1.png",
        "caption": "Fig. 4.11 Physical pendulum.",
        "texts": [
            " The purpose of this section is to study, by use of the analytical mechanics formalism and the theory developed in this chapter, both some classical applications and several extensions of this formalism to non-mechanical systems. This will be done by virtue of the analogies that can be identified between mechanical models and other physical representations. A rigid body able to oscillate about a fixed horizontal axis, which does not pass through the centre of mass, is a physical pendulum, or a compound pendulum (Fig. 4.11). Since the only applied force acting on the body is the force of gravity, the projection of the equation Mg \u00bc grad V on the axes of the frame S gives by integration the potential energy of the pendulum: Here, l \u00bc OG; M is the mass of the body, while the integration constant is determined from the condition: V = 0, for x = l. The instantaneous vector of rotation x is directed along the fixed axis and has the magnitude jxj \u00bc xz \u00bc _h: In view of (4.5.11), the kinetic energy is T \u00bc 1 2 I _h2; thus, the Lagrangian of the physical pendulum reads: L \u00bc 1 2 I _h2 Mgl\u00f01 cos h\u00de: \u00f04:6:1\u00de Let h be the generalized coordinate associated to the only degree of freedom of the body"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001541_978-3-642-35849-4_4-Figure4.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001541_978-3-642-35849-4_4-Figure4.2-1.png",
        "caption": "Fig. 4.2 Current cell deformed from reference orthogonal unit cell to which First Piola-Kirchhoff stress applies",
        "texts": [],
        "surrounding_texts": [
            "Designating the infinitesimal volumes in a specific region of material in the initial and the current configurations as dV and dv(= Jdv), respectively, the work rate \u2022 w0 done per the unit reference volume, i.e. a certain volume element possessing a fixed mass is given from Eqs. (2.40), (2.50), (2.80), (2.84), (2.117), (3.18), (3.24) and (3.28) as follows: \u2022 w0 = tr(\u03c3 d)dv/dV = tr(\u03c3 d)J = tr(\u03c4 l) = \u23a7 \u23aa \u23aa\u23aa\u23a8 \u23aa \u23aa\u23aa\u23a9 tr{(\u03c4 F\u2212T )(lF)T}= tr(\u03a0 \u2022 F T ) tr{(F\u22121\u03c4 F\u2212T )(FT dF)}= tr(S \u2022 E) = tr(S \u2022 C /2) tr(B \u2022 B) leading to \u2022 w0 = J\u03c3 : d = \u03c4 : d = \u03a0 : \u2022 F = S : \u2022 E = S : \u2022 C /2 =B : \u2022 B (4.87) where B \u2261 1 2 (SU+US) (4.88) which is called the Biot stress tensor. Eq. (4.87)7 is derived also as follows: tr(S \u2022 E) = tr [ S { 1 2 (U \u2022 U+ \u2022 U U) }] = 1 2 tr(SU \u2022 U+US \u2022 U) = 1 2 tr{(SU+US) \u2022 U}= tr(B \u2022 B) (4.89) noting Eqs. (2.40), (2.50)1 and (2.66). By taking account of Eq. (4.87)4 into the relation \u2022 w0 = 3 \u2211 i=1 fi \u2022 \u2022 n i = 3 \u2211 i=1 fi \u2022 (FNi) \u2022 = 3 \u2211 i=1 \u03a0Ni \u2022 (FNi) \u2022 = 3 \u2211 i=1 \u03a0Ni \u2022 \u2022 FNi = 3 \u2211 i=1 Ni \u2022 \u2022 F T \u03a0Ni = ( \u2022 F T \u03a0)ii = tr(\u03a0 \u2022 F T ) (4.90) we can confirm the fact that \u2022 w0 designates the work rate done in the current cell with the side vectors (n1,n2,n3) which was the orthogonal unit cell with the side vectors (N1,N2,N3) (||Ni||=1) in the reference state, noting Eqs. (1.103), (2.20) and (3.23) with the replacements dA\u2192 1 and df\u2192 f, as shown in Fig. 4.1. Besides, the first PiolaKircchoff stress is calculated supposing that the force fi on the current cell formed by the vectors (n1,n2,n3) applies to the reference dell formed by the vectors (N1,N2,N3). The work rate reflecting the constitutive property is not concerned with a current unit volume but is concerned with a reference unit volume, noting that the mass in the current unit volume is variable but the mass in the reference unit volume is invariable. The pairs of stresses and strain rates (or rates of deformation gradient) shown in Eq. (4.87) are called the work-conjugate pair. Stress and strain rate tensors in the workconjugacy pair have to be used for the formulation of constitutive equation."
        ]
    },
    {
        "image_filename": "designv11_100_0001878_9780470950029.ch12-Figure12.52-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001878_9780470950029.ch12-Figure12.52-1.png",
        "caption": "Figure 12.52 A turning pendulum.",
        "texts": [],
        "surrounding_texts": [
            "10. Quasi-Kinetic Energy Determine the following expressions:\n(a) N3 = 1\n3\n\u222b\nB\nv3dm\n(b) N4 = 1\n4\n\u222b\nB\nv4dm\n(c) N5 = 1\n5\n\u222b\nB\nv5dm\n(d) Is there a general rule to determine Nn based on N2 = K = 1 2 \u222b B v2dm?\n11. Falling Disc Assume the disc of Figure 12.51 has an initial velocity v0 at its center. Show that the velocity of the disc is\nvh = \u221a v2 0 + 2mgh\nm + I/R\nif it rolls without slip and drops a height h on a curved path. The disc has a mass m, radius R, and mass moment I .\nm x\ny\ng h\nR\nG",
            "13. Euler Equation for a Rigid Body with Nonuniform Density A rigid body B (Oxyz )\nis turning in a global coordinate frame G (OXYZ ) with angular velocity G\u03c9B . Determine the Euler equations of B if the body is a cube with size a \u00d7 a \u00d7 a and with the following densities:\n(a) \u03c1 = x2(a/2 \u2212 y)\n(b) \u03c1 = x2(a/2 \u2212 y)(a/2 \u2212 z)\n(c) \u03c1 = x2\n(a 2 \u2212 y )(a2 4 \u2212 z2 )\n14. Gyroscopic Effect A circular disc of mass m and mass moment I about its central axis rotates about its axis of symmetry with angular velocity . At the same time, the axis itself rotates in a horizontal plane with a precessional angular velocity \u03c9 as shown in Figure 12.53. A gravity moment mgl acts on the system. Find the value of \u03c9 compatible with this steady-state motion.\nX\n\u03c9\nG B\nl\nz\nx\nR mZ\n\u03a9\nY\ng\nFigure 12.53 A rotating disc about its axis of symmetry and about a varietal axis.",
            "15. A Turning Plate about a Symmetric Line on a Turntable Consider a uniform plate with mass m, side a, and mass moment [I ] that is mounted on a horizontal shaft as shown in Figure 12.54. The shaft is mounted on a table and is turning with constant angular velocity \u03b8\u03072 = \u03c9 with respect to the table. If the table is also turning with angular velocity \u03b8\u03071 = , then calculate:\n(a) How much force is supported by bearings of the shaft. (b) The moment 0M in the global frame.\nx0\nz0\n\u03b82\n\u03b81\nz2x1 B0\nB1\nB2 l\nz1\ny0\nx2\ny1\na\ny2\nC\nFigure 12.54 A turning square plate about a shaft on a turning table.\n16. A Turning Plate about the Centerline on a Turntable Change the orientation of the plate of Problem 15 as shown in Figure 12.55 and solve the problem again.\nz0\n\u03b82\n\u03b81\nz2x1\nB1\nB2\nl\nz1\nx2\ny1\na y2\nF\nF\nFigure 12.55 A turning plate about the centerline on a turntable."
        ]
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.68-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.68-1.png",
        "caption": "Fig. 2.68 4RPPaPa-type fully-parallel PM with coupled Sch\u00f6nflies motions defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 30, limb topology R||P\\Pa\\\\Pa",
        "texts": [
            " 15 48. 4PaPPaR (Fig. 2.60) Pa\\P\\\\Pa\\kR (Fig. 2.22u) Idem No. 4 49. 4PaRPaP (Fig. 2.61) Pa\\R\\Pa\\kR (Fig. 2.22v) Idem No. 34 50. 4PaRPPa (Fig. 2.62) Pa\\R||P\\Pa (Fig. 2.22w) Idem No. 34 51. 4PaPRPa (Fig. 2.63) Pa\\P||R\\Pa (Fig. 2.22x) Idem No. 33 52. 4PaPaPR (Fig. 2.64) Pa\\Pa\\\\P||R (Fig. 2.22y) Idem No. 4 53. 4PaPaRP (Fig. 2.65) Pa\\Pa\\\\R||P (Fig. 2.22z) Idem No. 41 54. 4RPaPaP (Fig. 2.66) R\\Pa\\\\Pa\\\\P (Fig. 2.22a0) Idem No. 15 55. 4RPaPPa (Fig. 2.67) R\\Pa\\kP\\\\Pa (Fig. 2.22b0) Idem No. 15 56. 4RPPaPa (Fig. 2.68) R||P\\Pa\\\\Pa (Fig. 2.22c0) Idem No. 15 57. 4PRPaPa (Fig. 2.69) P||R\\Pa\\\\Pa (Fig. 2.22d0) Idem No. 5 58. 4PaCPa (Fig. 2.70) Pa\\C\\Pa (Fig. 2.22e0) The axes of the cylindrical joints of the four limbs are parallel 59. 4CPaPa (Fig. 2.71) C\\Pa\\\\Pa (Fig. 2.22f 0) Idem No. 58 60. 4PaPaC (Fig. 2.72) Pa\\Pa\\\\C (Fig. 2.22g0) Idem No. 58 T ab le 2. 4 L im b to po lo gy an d co nn ec ti ng co nd it io ns of th e fu ll y- pa ra ll el so lu ti on s w it h no id le m ob il it ie s pr es en te d in F ig s. 2. 73 , 2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001858_978-94-007-7401-8_2-Figure2.2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001858_978-94-007-7401-8_2-Figure2.2-1.png",
        "caption": "Fig. 2.2 Fully-parallel PMs with coupled Sch\u00f6nflies motions of types 4RRRRR (a) and 4RRRRR (b) defined by MF \u00bc SF \u00bc 4; RF\u00f0 \u00de \u00bc \u00f0v1; v2; v3;xd\u00de, TF = 0, NF = 2, limb topology R\\R||R||R\\k R (a) and R||R\\R||R\\R (b)",
        "texts": [
            "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, 2.16, The limb topology and connecting conditions of the solutions in Figs. 2.2, 2.3, 2.20 are systematized in Table 2.1, as are their structural parameters in Table 2.2. Table 2.1 Limb topology and connecting conditions of the fully-parallel solutions with no idle mobilities presented in Figs. 2.2, 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, No. PM type Limb topology Connecting conditions 1. 4RRRRR (Fig. 2.2a) R\\R||R||R\\k R (Fig. 2.1a) The first and the last revolute joints of the four limbs have parallel axes 2. 4RRRRR (Fig. 2.2b) R||R\\R||R\\R (Fig. 2.1b) The first, the second and the last revolute joints of the four limbs have parallel axes 3. 4RRRRR (Fig. 2.3a) R||R||R\\R||R (Fig. 2.1c) The two last revolute joints of the four limbs have parallel axes 4. 4RRRRR (Fig. 2.3b) R||R||R\\R||R (Fig. 2.1c) The three first revolute joints of the four limbs have parallel axes 5. 4RRRRR (Fig. 2.4a) R||R\\R||R||R (Fig. 2.1d) The two first revolute joints of the four limbs have parallel axes. 6. 4PRRRR (Fig. 2.4b) P\\R\\R||R||R (Fig. 2.1e) The second joints of the four limbs have parallel axes 7"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0000955_978-3-030-73022-2_19-Figure31-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0000955_978-3-030-73022-2_19-Figure31-1.png",
        "caption": "Fig. 31 Heel chamfering with Chamfer Tool\u2014smooth chamfer edge",
        "texts": [
            " However, there is a potential issue with this type of tool in that if the diameter of the spherical end is small, which is required to fit in the fillet area, a double lip is created where the tool enters and exits the tooth edge, as Fig. 30 shows. In many instances, this is unacceptable and either an alternative solution is required, or else the BM diameter is increased to minimize the double lip and the fillet area is not targeted. Chamfer Tools (CT) are widely available and come in varied sizes. They offer an interesting, and inexpensive, alternative to End Mill tools since they can be used along tooth tips and, at least, at one tooth end, Toe or Heel. For example, the case of Fig. 30 would rather appear as shown in Fig. 31, left, with a result similar to what appears in Fig. 31, right: the quality of a CT tool chamfer is obvious. Seven unit vectors (Fig. 32) are required to orient the CT at any point along the tooth edge: N : the local normal vector, pivoted about \u2212\u2212\u2212\u2192 Pivot such as to face the cutting edge of the Chamfer Tool,\u2212\u2192 Vo: N X T ,\u2212\u2212\u2192 Tool: the tool vector,\u2212\u2212\u2212\u2192 Trans: axis about which vector \u2212\u2212\u2192 Tool is rotated by the Pivot Angle,\u2212\u2212\u2212\u2192 Pivot : axis about which vector \u2212\u2212\u2192 Tool is rotated by the local spiral angle,\u2212\u2212\u2212\u2212\u2212\u2192 VnCone: perpendicular to vector \u2212\u2212\u2192 Tool and the tooth edge and used to control the movement along the tool axis caused by Ds"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002586_msf.768-769.755-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002586_msf.768-769.755-Figure1-1.png",
        "caption": "Fig. 1 Fig. 2",
        "texts": [
            "s working on progressive products and technologies in modern research and development centers the mechanisms of rolling contact fatigue are studied since several decades [1-6]. The object of studies (Fig. 1) shows several sections of a distressed deep groove ball bearing inner ring. Metallographic evaluation, using Nital etchant, exhibits specific patterns so called \u201cdark etching areas (DEA)\u201d below the tread (Fig. 2). These DEA-patterns indicate plastic deformation of the microstructure, caused by overstraining. They arise during first revolutions and are going along with the whole fatigue procedure. These DEA-rims extend into deeper zones below the contact area (Fig. 3). Further structural alterations of the rolling contact fatigue process, so-called white etching areas (WEA), are then formed, as shown by means of metallography (Fig"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002764_0976-8580.99296-Figure7-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002764_0976-8580.99296-Figure7-1.png",
        "caption": "Figure 7: Angle at which the squeezed material runs off",
        "texts": [
            "6E 59.932 6.664 24.102 22.477 1.00 0.7E 22.374 5.96 23.411 21.268 1.50 0.8E 15.679 5.992 20.902 8.264 2.33 0.9E 18.196 5.774 7.392 6.649 4.00 From this table, it is clearly seen that the material less than E/Y=769.23, has the maximum stress value at a tangent modulus of 0.5E and above that the maximum stress value, at 0.3E Journal of Engineering and Technology | Jul-Dec 2012 | Vol 2 | Issue 2 101 The estimation of the angle at which the squeezed material comes out from the contact zone as shown in Figure 7 is important for contact analysis, to estimate the principle stresses in the sphere and a flat. The angle at which the squeezed material escapes is calculated and shown in Table 6. An estimation of the angle, as shown in Figure 8, at which the squeezed material comes out from the contact region provides a platform for analyzing the transformation of stress in the contact zone. In this finite element analysis the effect of the tangent modulus is studied. The result shows that an increase in the tangent modulus value increases the stress in the material (E/Y<1000) up to 0"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0001790_978-3-642-19853-3_78-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0001790_978-3-642-19853-3_78-Figure1-1.png",
        "caption": "Fig. 1. The geometric model of a 4SPS-2CCS parallel platform robot",
        "texts": [
            " Lazard, Ronga and Mourrain have proven that the general 6-6 hexapod FKP has 40 complex solutions using Greobner bases [1], Chern classes of vector bundles [2] and explicit elimination techniques [3]. Wen-Liang [4] and Zhang-Song [5] gave the closed-form solutions for the SP with planar base and platforms. For the general case, Lee [6] derived a set of 6th-dgree polynomial equations which lead to a 40th-degree univariate equation. In this paper, we propose a method to calculate the forward kinematics of parallel link manipulators. We calculate the forward kinematics numerically. The method is simple and can make the computer program easy to write. Fig. 1 shows the geometric model of a 4SPS-2CCS parallel platform robot. The absolute local frame system O1-X1Y1Z1 is fixed to the point A1 on the base platform and the relative moving frame system O2-X2Y2Z2 fixed to the point B1 on the moving platform. The direct kinematics problem is to find the position and orientation of the moving platform supposing that the pose of the base platform is known and values for the six constraints connecting to the base and the platform are given. Let Ai denote the coordinates of point Ai in the absolute frame O1-X1Y1Z1 and Bi denote the coordinates of point Bi in the moving frame O2-X2Y2Z2"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002680_amr.189-193.4173-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002680_amr.189-193.4173-Figure1-1.png",
        "caption": "Fig. 1 Hobbing principle",
        "texts": [
            " And a straight generatrix error may be regarded as the combination an involute error and a helix error. To some extent, straight generatrix errors are the comprehensive representation of profile and helix errors. Pathofcontact errors can be direct obtained by the method of comprehensive measuring integrated errors with worm single-tooth-face engagement, which is the typical application of measuring gear errors on the combination pathofcontacts and helicixes[3]. In the following, the shaping principle in hobbing involute gear tooth surfaces is studied on the basis of characteristic lines. In Fig.1a, the work-piece is rotated corresponding the hob According to the Transmission ratio, which is provided by the hobbing machine transmission chain. The hob can be moved relative to the work-piece along the work-piece axis, and there is an additional movement for hobbing helical gears, which are provided by differential motion train. And the hobbing principle is involute engagement principle with the differential motion. While having rotations of hob and work-piece, and having none of the axial movement and differential motion, a pathofcontact is hobbed on the gear tooth face. While having the axial movement and differential motion, and having none of the hob and work-piece rotations, a helix is marked on the gear tooth face, which is consist of 1Q , 2Q , 3Q \u2026 named as starting points in Fig.1b.The gear tooth face that is consist of pathofcontacts and helixes can be presented. While the start contact points of the hob and the work-piece are 1Q , ' 1Q ( in Fig.1b, only one tooth face can be presented, and ' 1Q is on the another tooth face that cannot be presented), Pathofcontact 1 can be hobbed with rotating of the hob and work-piece. After the work-piece is rotated a round, the contact point of the hob and work-piece is 2Q by mains of the axial movement and differential motion. Therefore Pathofcontact 2 can be hobbed. The tooth face can be hobbed with going on hobbing the other pathofcontacts. And we can understand that consistent relative positions of the hob and work-piece are presented at 1Q , 2Q , 3Q \u2026 "
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002737_09507116.2011.590664-Figure3-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002737_09507116.2011.590664-Figure3-1.png",
        "caption": "Figure 3. Measurements for the main dimensions of the bead.",
        "texts": [
            " In such cases, there is a risk that the melting of the groove face either in the groove or, in T-joint welding, on both sides of the penetration bottom may be incomplete, resulting in poor fusion. To confirm that appropriate weaving conditions can be set and adequate melting of the groove face is possible, a Tjoint of 6 mm thick SM490A steel sheet base metal was lap welded with an intended leg length of 6.5 mm, with the groove and torch settings as shown in Figure 2, and the penetration dimensions were measured as shown in Figure 3. Figure 4 shows a photograph of the appearance of the weldment and a macrograph of its cross section. It is clear from Figure 4 that fusion of the groove root area is unproblematic when welding is carried out with Ar \u00fe 3.5% O2 two-component mixed gas, with a penetration depth similar to that achieved when Ar \u00fe 3.5% O2 \u00fe 20% CO2 and with a good bead appearance. Figure 5 shows the relationship between spatter loss, shielding gas, and arc voltage when the carbon dioxide is 20%; if the arc voltage is set low to prevent undercut, there is an increase in the spatter quantity whereas with Ar \u00fe 3"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003396_amm.275-277.174-Figure1-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003396_amm.275-277.174-Figure1-1.png",
        "caption": "Fig. 1 The simplification of the rolling model Fig. 2 Load steps in rolling computation",
        "texts": [
            " Boundary Conditions and FE Model. It is not advisable to perform the three dimensional contact analysis with crankshaft model, because the irregular element shape and high density mesh in small surfaces would make the computation hard to converge or spend too much time. According to Saint Venant\u2019s principle, the crankshaft could be simplified to stepped shaft whose geometry character of fillet and journal is the same as crankshaft. The rolling model before and after simplification are shown in Fig. 1. Only half of the roller and stepped shaft model are considered in finite element computation. Note that the unit in Fig. 1 is millimeter. Roller and crankshaft are both revolving during fillet rolling process. The rotational speed is slow and the contact surface is quite smooth, so the inertia force and friction force are ignored in order to reduce the compute time consumption. Three steps are adopted to simulate the fillet rolling process as shown in Fig. 2. Section B is the symmetry plane of half stepped shaft model, which is also the dangerous section. In the first load step, the roller is put in the position of section I, and 0",
            " Results and Discussion. After the roller is unloaded, the stresses in plane B could be regarded as residual stress which induced by fillet rolling process. The most significant stress component for the fatigue analysis is the normal stress \u03c3\u03b8 in the polar coordinate system near fillet as shown in Fig. 4, which could be marked as \u03c3\u03b8r. The distributions of \u03c3\u03b8r in different depth a along \u03b8 are shown in Fig. 5. Boundary Conditions and FE Model. The geometry parameters of the crankshaft section not shown in Fig. 1 are list in Table 3. Because the geometry and bending loads are symmetrical, quarter model of crankshaft section is selected. The geometry model is divided into some blocks, and then hexahedron meshes are generated as shown in Fig. 6. In the fillet region of main journal and rod journal where stresses concentrate, local mesh refinement approach is used. Symmetry constraints are set in the symmetry plane of model, and displacements in vertical direction are restrained by weak spring. After that, static bending moment of 2400 Nm is applied to the model"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003409_2013-01-2780-Figure8-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003409_2013-01-2780-Figure8-1.png",
        "caption": "Figure 8. RSFB region",
        "texts": [
            " Boundary conditions for the identified critical regions are captured from super element without any user intervention. Figure 6 and Figure 7 shows the comparison of displacements and stresses respectively without and with sub-modeling. Displacement and stress results with sub-modeling is compared with full model linear static analysis and observed the same results. This has given confidence that submodeling is properly done. FE model should be modeled without any penetration. This is a pre-requisite to carry out contact analysis otherwise it will lead to unrealistic results. Figure 8 shows the super element at RSFB region and Figure 9 shows the Penetration plot at RSFB region. It is observed that bottom region of the suspension mounting bracket is penetrating inside the external flitch. Figure 10 shows the comparison of stresses with and without penetration at RSFB region. It is observed from Figure 10 that unrealistic stress occurs at the location where penetration happens. Table 1 shows the comparison of stresses and displacements with and without penetration. It is observed from Table 1 that there is a significant difference in stresses"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002792_12.876079-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002792_12.876079-Figure2-1.png",
        "caption": "Figure 2. Schematic of the optical system applied: 1- fiber laser; 2- beam expander; 3- laser beam /thermal signal separating mirror; 4 - scanner head; 5 - F-Theta lens; 6 - powder bed; 7 - mirror; 8 - pyrometer lens; 9 - fiber tip; 10 \u2013 optical fiber; 11- pyrometer.",
        "texts": [
            " To exclude the influence of the scanning speed, observations must be carried out coaxially to the laser beam. In this case, for any position of the laser beam and for any scanning speed, the diagnostic system will register the signal from the processing zone. At the given state of technical knowledge in this field, pyrometry has been chosen among other optical methods. In order to continuously monitor the surface temperature, i.e. temperature in the laser impact zone, a two-wavelength pyrometer (11, fig. 2) registering surface thermal radiation is connected to the optical unit by optical fiber (10, fig. 2). Temperature was measured in the laser impact zone by an originally developed bi-colour pyrometer with the following performance: temperature range is 900-2600\u00baC, two InGaAs photodiodes with optical filters, transmission spectrum is 1.26\u00b5m and 1.4\u00b5m, respectively, with 100nm bandwidth, 560\u00b5m diameter of the zone of temperature measurements, 50 ms sampling time. At the present time, arbitrary units are used instead of the temperature values because of two reasons. The first one is the size of the pyrometer view field"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0002714_amm.275-277.736-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0002714_amm.275-277.736-Figure2-1.png",
        "caption": "Fig. 2. Border conditions and load demonstrations of calculation conditions of lifter",
        "texts": [
            " As the equipment parts are large, many additional calculations are needed if solid contact chosen. To make calculation easier, the calculation will become more reasonable if defined surface shell elements contact chosen. Load working conditions and border conditions According to the actual station, there are 8 types of calculation working conditions as per various stations and loading conditions, including simulation load working condition. Refer to Table 1 for the load working conditions and border conditions in details. Fig. 2 Displacement restraints of simulated load working conditions and the location and values of all load conditions on lifter are given. Calculation result of static strength The calculation result for each working condition can be read through the post processing function of ANSYS so as to get deformation nephogram and stress nephogram of various working conditions; the location and the stress value of the second largest stress point can be obtained by deleting the largest stress point. 5 stress points shall be chosen for each working condition in order to find as many as possible the possible failure locations"
        ],
        "surrounding_texts": []
    },
    {
        "image_filename": "designv11_100_0003298_amr.179-180.150-Figure2-1.png",
        "original_path": "designv11-100/openalex_figure/designv11_100_0003298_amr.179-180.150-Figure2-1.png",
        "caption": "Figure 2: Mathematic model for yarn length change",
        "texts": [
            " Due to more yarn consumption, the yarn tension will become higher than needle bar not shifts, which causes more rebound length and withdrawal value therefore producing lower pile than predefined pile height. If the needle bar shifts while tufting high piles but not in tufting low piles, then the height error will increase. The needle bar transverse will affect both in needle tracking and yarn guide route, which generates additional yarn consumption represented with L\u2206 . The additional usage of yarn length due to varying needle tracking is marked as tL\u2206 and the additional length due to varying yarn guide route is marked as yL\u2206 , and get L\u2206 as Eq. 1. t y= +L L L\u2206 \u2206 \u2206 (1) As Fig. 2 shown is the yarn usage in rectangular coordinate. Before feeding the yarn to individual needles, it needs to cross over two sets of thread-carriers from the yarn feed roller. One fixed guide plate as 10 in Fig.1 is fixed on the machine frame, which facilitates the guiding of the yarn onto the yarn feed roller. Another two guide plates are oscillatory with needle bar. The alternation of relative place for switching between the two thread-carriers is used to store yarn and adjust yarn tension. As shown in Fig. 2, consider the yarn as a rigid body and set up the rectangular coordinate, the grid origin is the fix outlet terminal of fix guide carrier. The positive y-axis is to direct the motion of the backing cloth. It can be rotated 90 degrees clockwise onto the x-axis. The positive z-axis is known from the right-hand screw rule. Point nP in the coordinate represents the hole on the float thread-carrier in which pass through the same yarn as fix hole O on fixed thread-carrier. A0 is the projective point of needle at the backing cloth when the needle bar does not shift",
            " The above expression shows that the compensation of the left traverse (from point n+1P to point nP or B1 to A2) is less than the right traverse (from point nP to point n+1P or A0 to B1). In Plain terms extra yarn feed amounts for traverse left and right is different. In Eq. 7 ntg /S b\u03b2 = and n+1tg ' /S b\u03b2 = 2 2 2 2 2 2 y n n+1=L b S h b S h\u2206 + + \u2212 + + (1) As n+1 tr nS L S= + we could get L\u2206 as (10): 22 2 2 2 22 2 st tr nnst tr = + ( )L L L Sb S h b hL L+ \u2212 + + \u2212 + + +\u2206 (10) As Eq.5 shown the more yarn usage is determined by transverse shifting distance trL and nS . The more trL will lead to more yarn usage on the carpet backing as Fig.2 A0B1 is longer than original length stL . The yarn feeding path change is determined by nS and n+1S , if n n+1<S S , the yarn feeding path will cost more yarn, otherwise the feeding path become shorter and would provide redundant length yarn for loop forming. To calculate L\u2206 , we need to know the value of nS and trL (or n+1S ). nS and trL is concern with the needle bar shifting track. In our developed carpet tufting machines, parameter b value is 155.00 mm, and h is 74.90 mm, stitch rate is 44.0 stiches/10cm, the range of motion of needle is 31 mm"
        ],
        "surrounding_texts": []
    }
]