diff --git "a/designv11-3.json" "b/designv11-3.json" new file mode 100644--- /dev/null +++ "b/designv11-3.json" @@ -0,0 +1,9318 @@ +[ + { + "image_filename": "designv11_3_0002637_s10846-013-9827-5-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002637_s10846-013-9827-5-Figure1-1.png", + "caption": "Fig. 1 Two-link manipulator model", + "texts": [ + " In the simulation part, we show that even though the controller has a marginal performance, the neural network desirably works. The rest of the paper is as follow. In the next section the model of the system is presented. The multivariable feedback linearization is proposed in Section 3. In Section 4, the neural network is considered and learning rules are proposed. The robust controller is designed in Section 5. The application and performance of the proposed controller on the system are studied in Section 6. Finally, the conclusion is presented in Section 7. The model of the system is depicted in Fig. 1. The equation of the involved model is described as follow [1\u20134]: M(\u03b8)\u03b8\u0308 + N(\u03b8, \u03b8\u0307 ) + H(\u03b8) = \u03c4 (1) where \u03b8 = [\u03b81, \u03b82]T is the output vector, \u03c4 = [\u03c41, \u03c42]T is the input torque vector, M(\u03b8) = [M11, M12; M21, M22] is the symmetric inertia matrix that means M12 = MT 21. Forasmuch as we consider the two-link manipulator, M12 and M21 are scalar values that result M12 = M21. The elements of M(\u03b8), N(\u03b8, \u03b8\u0307 ) = [N1, N2]T , H(\u03b8) = [H1, H2]T matrices are gained as follow [1\u20134, 19]: M11 = m1l2 1 3 + m2l2 2 12 + m2 ( l2 1 + l2 2 4 ) + ml ( l2 1 + l2 2 ) + (m2 + 2ml) l1l2 cos \u03b82 M12 = m2l2 2 12 + m2l2 2 4 + mll2 2 + (m2 + 2ml)l1l2 cos \u03b82 2 M22 = m2l2 2 12 + m2l2 2 4 + mll2 2 N1 = \u2212(m2 + 2ml)l1l2\u03b8\u03071\u03b8\u03072 sin \u03b82 \u2212 (m2 + 2ml)l1l2\u03b8\u0307 2 2 sin \u03b82 2 N2 = (m2 2 + ml ) l1l2\u03b8\u03071 sin \u03b82 H1 = ( 1 2 m1l1 + m2l2 + mll1 ) g sin \u03b81 + ( 1 2 m2l2 + mll2 ) g sin (\u03b81 + \u03b82) H2 = ( 1 2 m2 + ml ) l2g sin (\u03b81 + \u03b82) where m1, m2, ml, l1, l2, g are the mass of the first link, the second link and the load, length of first link and second link, and Earth\u2019s gravity, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002314_icsse.2011.5961870-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002314_icsse.2011.5961870-Figure4-1.png", + "caption": "Figure 4. Gearbox fault simulator", + "texts": [ + " Step3: Adjust weighted value: According the error value ek to update value of weighted and threshold are made. The weighted value update: 1 (1 ) ( ) M ij ij j j jk k k w w H H x i w e\u03b7 = = + \u2212 (5) jk jk j kw w H e\u03b7= + (6) The threshold value update: 1 (1 ) M j j j j jk k k a a H H w e\u03b7 = = + \u2212 (7) k k kb b e= + (8) Where represents the learning rate and the available value is situated between 0.1 to1. By repeating the above three steps, the error value ek will be zero or fixed small value. The gearbox faults were simulated by a test rig shown in Fig. 4, which can simulate most faults than can commonly occur in gearbox, such as unbalance, misalignment, gear crack, and so on. In this study, a total 9 classes conditions were analyzed which included the failure of gear, and structural faults. The structural faults such as unbalance, misalignment and looseness were analyzed. The five types of gear faults are used in experiment with gear tooth broken, chipped tooth, gear crack, gear tooth broken combine with chipped tooth and gear crack combine with chipped tooth, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001190_icelmach.2010.5608143-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001190_icelmach.2010.5608143-Figure10-1.png", + "caption": "Fig. 10 Photograph of torque measuring system.", + "texts": [ + " This implies the flux path helps to reduce the cogging torque on the high-speed rotor. The orders shown in Table I are found. The orders not shown there are the harmonic components (H.C.) by its pole pairs and stationary pole pieces. These orders appear because of the error of the mesh in the analysis model. VI. VERIFICATION BY A PROTOTYPE A prototype based on the specifications mentioned above was manufactured, and is shown in Fig. 9. This prototype is set to the torque measuring system shown in Fig. 10, and the constant rotation speed was given to both rotors in accordance with the gear ratio by the AC servo motors. A rotation speed of 0.5 rpm was given to the high-speed rotor to ignore the effect of the eddy current. The synchronous transmission torque shown in Fig. 11 was measured. The orders contained in the cogging torque on both rotors are shown in Figs. 12 and 13. The cogging torque on the high-speed rotor and lowspeed rotor is 0.064 Nm and 0.371 Nm, respectively, and higher than the computed values" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure6-1.png", + "caption": "Fig. 6 Illustration of: \u201ea\u2026 pinion and gear root lines and axes zP and zF of corresponding generating surfaces, \u201eb\u2026 gear generator and gear pitch cones, and \u201ec\u2026 pinion generator and pinion pitch cones", + "texts": [ + " Nevertheless, the generation of pinion and gear tooth surfaces y the same generating surface is not usually applied since the earing contact needs to be localized. For this purpose, different enerating surfaces are applied for pinion and gear. Derivation of Gear Machine-Tool Settings The following subsections show the derivations of gear achine-tool settings for the case of a spiral bevel gear drive with tandard taper. The obtained relations are then extended for the ase of uniform and duplex taper. 4.1 Derivation for Standard Type of Taper. Figure 6 a hows the pinion and gear root lines, which are not parallel but ntersect each other at the common apex of pitch cones, point O. ere, axis zf coincides with axis zcg and axis xf coincides with the nstantaneous axis of rotation, line OI see Fig. 4 a . Since the xis of rotation of the generating surface must be perpendicular to he closer generatrix of the root cone, different axes of rotation of he generating surfaces are needed for generation of pinion and ear. Axis zP is perpendicular to the gear root line whereas axis zF s perpendicular to the pinion root line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002754_mmar.2013.6670008-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002754_mmar.2013.6670008-Figure2-1.png", + "caption": "Fig. 2. Kinematical structure of the drive chain with gear boxes and disturbance torques.", + "texts": [ + " As a result, the differential equation for the difference pressure is given by \u2206p\u0307 = 2 CH ( V\u0303P \u03b1\u0303P\u03c9P \u2212 V\u0303M \u03b1\u0303M\u03c9M \u2212 qU 2 ) , (10) with a lumped disturbance in form of the resulting leakage volume flow qU = 2kIL\u2206p+ kELApA \u2212 kELBpB . (11) The longitudinal dynamics of the working machine is governed by the equation of motion. The vehicle with the drive chain system (vehicle mass mv , wheel radius rw, gear box transmission ratio ig , rear axle transmission ratio ia, damping coefficient dg at the drive shaft), see also Fig. 2, can be described by the following second order differential equation( JM + Jg ig 2 + Ja +mV r 2 w i2a i 2 g\ufe38 \ufe37\ufe37 \ufe38 JV ) \u03c9\u0307M + dg i2g\ufe38\ufe37\ufe37\ufe38 dV \u03c9M = (12) V\u0303M\u2206p \u03b1\u0303M\ufe38 \ufe37\ufe37 \ufe38 \u03c4M \u2212 ( \u03c4Mf tanh( \u03c9M \u03b5 ) + \u03c4gf tanh( \u03c9M ig\u03b5 ) + \u03c4L iaig\ufe38 \ufe37\ufe37 \ufe38 \u03c4U ) , where \u03c4M is the torque of the hydraulic motor. JM , Jg and Ja are the mass moments of inertia of the hydraulic motor, gear box and rear axle, respectively. The maximum values \u03c4Mf and \u03c4gf characterise the friction models of the hydraulic motor and the gear box, whereas \u03b5 << 1 represents a small number" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003879_j.ymssp.2013.04.009-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003879_j.ymssp.2013.04.009-Figure4-1.png", + "caption": "Fig. 4. Planet motion in the (a) rotating carrier basis and (b) stationary basis when vibrating in a single planet mode at \u03a9c \u00bc 0:5000 (Point J in Fig. 2). In (b) the motions of planets 2, 3, and 4 are not shown. The motion for planet 1 is shown for 13 periods of the natural frequency. The circle markers denote the initial planet positions.", + "texts": [ + " From Eq. (15), it can be shown that if \u03b6 and \u03b7 have the same sign the planets orbit about their nominal equilibrium in the direction opposite of carrier rotation. When the signs of \u03b6 and \u03b7 differ the planets orbit in the direction of carrier rotation. Transforming Eq. (15) to the stationary basis gives ri \u00bc fR cos\u00f0\u03a9ct \u00fe \u03c8 i\u00de \u00fe jwij\u00f0\u03b6 \u00fe \u03b7\u00de cos\u00bd\u00f0\u03c9\u00fe\u03a9c\u00det \u00fe \u03b3i \u00fe \u03c8 i \u00fe jwij\u00f0\u03b6\u2212\u03b7\u00de cos\u00bd\u00f0\u03c9\u2212\u03a9c\u00det \u00fe \u03b3i\u2212\u03c8 i gE1 \u00fefR sin\u00f0\u03a9ct \u00fe \u03c8 i\u00de \u00fe jwij\u00f0\u03b6\u2212\u03b7\u00de sin\u00bd\u00f0\u03c9\u00fe\u03a9c\u00det \u00fe \u03b3i \u00fe \u03c8 i \u2212jwij\u00f0\u03b6 \u00fe \u03b7\u00de sin\u00bd\u00f0\u03c9\u2212\u03a9c\u00det \u00fe \u03b3i\u2212\u03c8 i gE2: \u00f016\u00de Fig. 4 shows the motion of the planet center while oscillating in a single planet mode at \u03a9c \u00bc 0:5000. The natural frequency is \u03c9\u00bc 0:8123. This is point J in Fig. 2. Fig. 4a shows the motion of each planet in the rotating carrier basis, calculated from Eq. (15). The planets move in elliptical orbits with the semi-major and semi-minor axes aligning with ei1 and ei2. All planet motions have the same foci because jwij \u00bc 1 for systems with four equally spaced planets. The phase difference can be inferred from the initial positions of each planet (denoted by circle markers). For four-planet systems with equally spaced planets, opposing planets are in-phase and adjacent planets are 1801 out-of-phase. Fig. 4b shows the motion of the center of planet 1 in the stationary basis calculated from Eq. (16). The motions of planets 2, 3, and 4 are not shown. The motion of planet 1 has more complicated geometry in the stationary basis due to the response having multiple frequency components, including \u03a9c \u00bc 0:5000, \u03c9\u2212\u03a9c \u00bc 0:3123, and \u03c9\u00fe\u03a9c \u00bc 1:3123. The response is shown for 13 periods of the natural frequency. It is nearly periodic because this gives nearly eight periods of \u03a9c, five periods of \u03c9\u2212\u03a9c, and 21 periods of \u03c9\u00fe\u03a9c", + " 6c shows the motion of planet 1 in the stationary frame at \u03c9\u00bc\u03a9c. Figs. 6d and e show the planet motion when the natural frequency is an integer fraction of carrier speed. In this case the planet motion has multiple loops, with the number of loops equal to the ratio of carrier speed to natural frequency. The curves traced by the planet motions in Figs. 6c and d are similar to lima\u00e7ons. For all cases the motions of each planet in the rotating carrier basis are elliptical orbits, as was the case shown in Fig. 4a at \u03a9c \u00bc 0:5000. Planetary gears have six rotational modes where the central members have only rotation and no translation. Additionally, each planet has identical modal displacements. A representative rotational mode is shown in Fig. 3b. Rotational modes have the form \u03d5\u00bc \u00bdpc;pr ;ps;p1;\u2026;p1 T , where ph \u00bc \u00bd0;0;uh T for h\u00bc c; r; s [21]. For single-mode vibration, each planet has identical motion relative to its local reference frame calculated from Eq. (9b) as ri \u00bc \u00bdR\u00fe 2j\u03b61j cos\u00f0\u03c9t \u00fe \u03b3\u03b6\u00de ei1 \u00fe \u00bd2j\u03b71j cos\u00f0\u03c9t \u00fe \u03b3\u03b7\u00de ei2; \u00f017\u00de where tan\u03b3\u03b6 \u00bc Im\u00f0\u03b61\u00de=Re\u00f0\u03b61\u00de and tan \u03b3\u03b7 \u00bc Im\u00f0\u03b71\u00de=Re\u00f0\u03b71\u00de" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001937_icems.2011.6073463-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001937_icems.2011.6073463-Figure2-1.png", + "caption": "Fig. 2. Schematic diagram of the 6-mass drive train modeling", + "texts": [ + " On the other hand, the vector control approach is used with a reference frame oriented along the stator flux linkage vector position on the rotor-side converter, in order to implement the active and reactive power decoupling control of DFIG. As a result, the electromagnetic torque of DFIG TG is given as: ( / )G m s P s rdT L L n i\u03c8= \u2212 (2) Where Lm is the inductance between rotor and stator, Ls is the inductance of stator, nP is the pair of poles, \u03c8s is the stator flux linkage, ird is the d component of rotor current. The drive train modeling of wind turbines is the emphasis in this paper, and the detailed modeling of each mass model is discussed next. A. 6-Mass Drive Train Modeling The 6-mass model shown in Fig. 2 has six inertias, namely three blade inertias (HB1, HB2 and HB3), hub inertia (HH), gearbox inertia (HGB), and generator inertia (HG). The elasticity between adjacent masses is expressed by the spring constants KBH1, KBH2, KBH3, KLS and KHS (LS and HS represent low- and high- speed shaft respectively). The mutual-damping between contiguous masses is expressed by DBH1, DBH2, DBH3, DLS and DHS. There exist some torque losses through external damping elements of individual masses, which are represented by DB1, DB2, DB3, DH, DGB and DG" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003469_j.oceaneng.2013.01.001-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003469_j.oceaneng.2013.01.001-Figure6-1.png", + "caption": "Fig. 6. Example of 4-legged robot system.", + "texts": [ + " In this paper, the simulations have been performed with four cases: (1) 4-legged robot considering differential pose, (2) 4-legged robot considering differential environment, (3) 4-legged robot considering differential tidal current, (4) 6-legged robot considering differential environment. In the simulation, we have applied the proposed method to the case without underwater environment condition for considering an influence of only leg\u2019s pose. Figs. 7 and 8 show the simplified models of the 4-legged robot with 3 D.O.F as shown in Fig. 6. In these models, one\u2019s legs are folded and the other\u2019s legs are spread. The parameters of each case are defined as Table 2. Here, the friction angle is selected as ym \u00bc 303 arbitrarily. Before analysis of its simulation result, we can easily expect several results from the configuration of robot: (1) the translational acceleration bounds along X and Y axes are symmetric, (2) the magnitude of acceleration along positive Z axis is larger than one along negative Z axis, (3) the rotational acceleration bounds along X, Y, and Z axes are symmetric, (4) in the translation along Y and Z axes, the acceleration bounds of folded leg\u2019s case are bigger than ones of spread leg\u2019s case, (5) in the rotation along X, Y, and Z axes, the acceleration bounds of folded leg\u2019s case are bigger than ones of spread leg\u2019s case" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002486_j.cnsns.2013.08.018-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002486_j.cnsns.2013.08.018-Figure2-1.png", + "caption": "Fig. 2. Typical rotor configuration and the fixed and rotating coordinates.", + "texts": [ + " The configuration of the geared rotor-bearing system with a slant crack on the shaft is shown in Fig. 1. Two uniform flexible shafts are of the same length L. The slant crack would be found on one of the two shafts, and its position is denoted by xc . The gear pair is modeled as two rigid disks mounted at a distance x1. The bearings are modeled as flexible elements with damping and stiffness coefficients denoted as cb and kb, respectively. A fixed reference frame, X\u2014Y\u2014Z, is used to describe the system motion. A single shaft system with rigid disk in the Y\u2014Z plane is shown in Fig. 2. Five degrees of freedom V ;W;a;B;C are considered at each nodal point of the shaft, in which the V and W are lateral displacements in Y\u2014Z directions, and B;C are rotation angles about Y\u2014Z directions. The torsional displacement is denoted by a and axial translational vibration is neglected. In order for deriving the flexibility matrix of the slant crack, a rotating reference frame x\u2014y\u2014z is also introduced as shown in Fig. 2. Frame x\u2014y\u2014z rotates at s speed of X about the X-axis with X denoting the spin speed of the rotor. The components of the system include disks, gear pair, bearing supports, rotor shafts and the slant crack. The equations of motion are derived for each component as follows. The kinetic energy of a disk with eccentricity for lateral motion given by Shiau and Hwang [26] is modified to include the torsional kinetic energy Td \u00bc 1 2 md V2 d \u00feW2 d \u00fe 1 2 IDd _B2 d \u00fe _C2 d \u00fe 1 2 IPd\u00f0X\u00fe _ad\u00de _BdCd _CdBd \u00femede\u00f0X\u00fe _ad\u00de2 _Vd sin\u00f0Xt \u00fe ad \u00feud\u00de \u00fe _Wd cos\u00f0Xt \u00fe ad \u00feud\u00de \u00fe 1 2 mede2\u00f0X\u00fe _ad\u00de2 \u00fe 1 2 IPd\u00f0X\u00fe _ad\u00de2 \u00f01\u00de where Vd and Wd are the two lateral displacements along the Y and Z directions and Bd;Cd are the corresponding bending angles in the Y\u2014Z plane, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002486_j.cnsns.2013.08.018-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002486_j.cnsns.2013.08.018-Figure4-1.png", + "caption": "Fig. 4. Slant cracked shaft finite element: (a) the element showing forces acting and coordinate; (b) top view of the shaft element with slant crack showing position of the crack relative to shaft axis; (c) crack cross-section for computation.", + "texts": [ + " Thus, the corresponding force model is expressed as [27] Fby Fbz Mbz Mby 2 6664 3 7775 \u00bc kby 0 0 0 0 kbz 0 0 0 0 0 0 0 0 0 0 2 6664 3 7775 Vb Wb Bb Cb 2 6664 3 7775 cby 0 0 0 0 cbz 0 0 0 0 0 0 0 0 0 0 2 6664 3 7775 _Vb _Wb _Bb _Cb 2 6664 3 7775 \u00f013\u00de in which the first matrix is the element stiffness matrix and the second is the viscous damping matrix. These matrices are diagonal and symmetric. Consider a rotor segment of radius R and length l with a slant crack having a depth of a oriented at an angle of h relative to the axis of the shaft, as shown in Fig. 4. The dimensionless crack depth are represented as r \u00bc a=R. The element is loaded with shear forces P1; P2 and P6; P7, bending moments P3; P4 and P8; P9 and torsional moments P5 and P10 (Fig. 4(a)). The center of the crack is situated at a distance 0:5l from the left end of the element as shown in the figure. The flexibility matrix of the cracked shaft element is derived. According to the Castigliano theorem, the flexibility coefficients of the element corresponding to the shear forces, bending moments, torsional moment and the combinations of both are expressed as fij \u00bc @ @Pi@Pj Uo \u00fe Uc\u00f0 \u00de \u00f0i; j \u00bc 1;2; . . . 5\u00de \u00f014\u00de where Uo is the strain energy of the uncracked shaft element, and Uc the additional strain energy due to the slant crack", + " \u00f015\u00de The additional strain energy due to the presence of slant crack is expressed as follows Uc \u00bc 1 E0 Z Z Ac X5 i\u00bc1 KI i !2 \u00fe X5 i\u00bc1 KII i !2 \u00fems X5 i\u00bc1 KIII i !2 0 @ 1 AdAc \u00f016\u00de in which E0 \u00bc E=\u00f01 m2\u00de and ms \u00bc 1\u00fe m;KI i ;K II i and KIII i represent SIFs corresponding to the opening, sliding and tearing modes of crack displacement, respectively. The forces P1\u2014P10 at the nodes act in a coordinate X\u2014Y\u2014Z aligned along the shaft axis. The stresses due to these forces are resolved in another coordinate system aligned with the slant crack x0\u2014y0\u2014z0, as shown in Fig. 4(a) and (c). Darpe [21] presented the SIFs for the three modes, and the flexibility coefficients fij could be obtained by solving a series of double integration. The detailed process is given in Appendix A. In the following, unless otherwise specified, the slant crack is referred to as crack with h \u00bc p=4. According to the crack geometry in Fig. 4(b), one can find that the slant crack would become a straight crack by setting h \u00bc p=2. Thus, the total flexibility matrix for a general orientation of crack is given as f \u00bc f11 f12 f13 f14 f15 f22 f23 f24 f25 f33 f34 f35 Sym: f44 f45 f55 2 6666664 3 7777775 \u00f017\u00de The above flexibility matrix is full populated as against the flexibility matrix of the straight crack, wherein some of the cross-coupled flexibilities do not exist, e.g. f35 \u00bc f45 \u00bc 0. These cross-flexibility coefficients are important as they indicate the direct coupling of torsional mode (i \u00bc 5) with the lateral modes (i \u00bc 3;4)", + " SIFs for Mode I (opening mode): KI 1 \u00bc jP1 pR2 sin 2h ffiffiffiffiffiffiffi pl p F1 KI 3 \u00bc 4\u00f0P2\u00f00:5l\u00fe b cos h\u00de P3\u00de pR4 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R2 b2 sin2 h q sin2 h ffiffiffiffiffiffiffi pl p F2 KI 4 \u00bc 4\u00f0P1\u00f00:5l\u00fe b cos h\u00de \u00fe P4\u00de pR4 b sin3 h ffiffiffiffiffiffiffi pl p F1 KI 5 \u00bc 2P5 pR4 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R2 b2 sin2 h q sin 2h ffiffiffiffiffiffiffi pl p F2 KI 2 \u00bc 0 \u00f0A:1\u00de SIFs for Mode II (sliding mode): KII 2 \u00bc jP2 pR2 sin h ffiffiffiffiffiffiffi pl p FII KII 5 \u00bc 2P5 pR4 b sin2 h ffiffiffiffiffiffiffi pl p FII KII 1 \u00bc KII 3 \u00bc KII 4 \u00bc 0 \u00f0A:2\u00de SIFs for Mode III (tearing mode): KIII 1 \u00bc jP1 pR2 cos 2h ffiffiffiffiffiffiffi pl p FIII KIII 3 \u00bc 4\u00f0P2\u00f00:5l\u00fe b cos h\u00de P3\u00de pR4 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R2 b2 sin2 h q sin h cos h ffiffiffiffiffiffiffi pl p FIII KIII 4 \u00bc 4\u00f0P1\u00f00:5l\u00fe b cos h\u00de \u00fe P4\u00de pR4 b sin2 h cos h ffiffiffiffiffiffiffi pl p FIII KIII 5 \u00bc 2P5 pR4 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R2 b2 sin2 h q cos 2h ffiffiffiffiffiffiffi pl p FIII KIII 2 \u00bc 0 \u00f0A:3\u00de where l is the depth of crack at any distance b from the center along the crack edge (axis z0 in Fig. 4(c)), and F1\u00f0l=l0\u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2l0 pl tan pl 2l0 s 0:752\u00fe 2:02\u00f0l=l0\u00de \u00fe 0:37 1 sin pl 2l0 h i3 cos pl 2l0 F2\u00f0l=l0\u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2l0 pl tan pl 2l0 s 0:923\u00fe 0:199 1 sin pl 2l0 h i4 cos pl 2l0 FII\u00f0l=l0\u00de \u00bc 1:122 0:561\u00f0l=l0\u00de \u00fe 0:085\u00f0l=l0\u00de2 \u00fe 0:18\u00f0l=l0\u00de3ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 \u00f0l=l0\u00de p FIII\u00f0l=l0\u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2l0 pl tan pl 2l0 s \u00f0A:4\u00de in which l0 is the total height of the strip of width db (Fig. 4(c)). With the expressions of SIFs given above, the additional strain energy due to crack could be estimated using Eq. (16). In order to facilitate the derivation, two variables are written as: xs \u00bc 0:5l\u00fe b cos h and R2 b \u00bc R2 b2 sin2 h. Some non-dimensional parameters are introduced: b \u00bc b=R; l \u00bc l=R; xs \u00bc xs=R and Rb \u00bc Rb=R. Thus, the expressions of flexibility for the slant cracked shaft are derived utilizing the total strain energy as follows [21] f11 \u00bc l jGA \u00fe l3 3EI \u00fe 1 E0pR Z b b Z a 0 ( 32 x2 s b2 l\u00f0sin6 hF2 1 \u00fems sin4 h cos2 hF2 III\u00de \u00fe 16 xsj b l\u00f0sin3 h sin 2hF2 1 \u00fems sin2 h cos h cos 2hF2 III\u00de \u00fe 2j2 l\u00f0F2 1 sin2 2h\u00fems cos2 2hF2 III\u00de ) d ld b \u00f0A:5\u00de f22 \u00bc l jGA \u00fe l3 3EI \u00fe 1 E0pR Z b b Z a 0 32 x2 s R2 b l sin4 hF2 2 \u00fe 2j2 l sin2 hF2 II \u00fe 32ms x2 s R2 b l sin2 h cos2 hF2 III n o d ld b \u00f0A:6\u00de f33 \u00bc l EI \u00fe 32 E0pR3 Z b b Z a 0 R2 b l sin4 hF2 2 \u00fems sin2 h cos2 hF2 III d ld b \u00f0A:7\u00de f44 \u00bc l EI \u00fe 32 E0pR3 Z b b Z a 0 b2 l sin6 hF2 1 \u00fems sin4 h cos2 hF2 III d ld b \u00f0A:8\u00de f55 \u00bc l GJs \u00fe 8 E0pR3 Z b b Z a 0 l R2 b sin2 2hF2 2 \u00fe b2 sin4 hF2 II \u00femsR2 b cos2 2hF2 III n o d ld b \u00f0A:9\u00de f12 \u00bc 32 E0pR Z b b Z a 0 x2 s Rb b l sin5 hF1F2 \u00fems sin3 h cos2 hF2 III d ld b\u00fe 8 E0pR Z b b Z a 0 xsRbj l sin2 h sin 2hF1F2 \u00fems sin h cos h cos 2hF2 III d ld b \u00f0A:10\u00de f13 \u00bc 32 E0pR2 Z b b Z a 0 xsRb b l sin5 hF1F2 \u00fems sin3 h cos2 hF2 III d ld b 8 E0pR2 Z b b Z a 0 Rbj l sin2 h sin 2hF1F2 \u00fems sin h cos h cos 2hF2 III d ld b \u00f0A:11\u00de f14 \u00bc l2 2EI \u00fe 32 E0pR2 Z b b Z a 0 xs b2 l sin6 hF2 1 \u00fems sin4 h cos2 hF2 III d ld b\u00fe 8 E0pR2 Z b b Z a 0 j b l sin3 h sin 2hF2 1 \u00fems sin2 h cos h cos 2hF2 III d ld b \u00f0A:12\u00de f15 \u00bc 16 E0pR2 Z b b Z a 0 xsRb b l sin3 h sin 2hF1F2 \u00fems sin2 h cos h cos 2hF2 III d ld b\u00fe 4 E0pR2 Z b b Z a 0 Rbj l sin2 2hF1F2 \u00fems cos2 2hF2 III d ld b \u00f0A:13\u00de f23 \u00bc l2 2EI 32 E0pR2 Z b b Z a 0 xsR2 b l sin4 hF2 2 \u00fems sin2 h cos2 hF2 III d ld b \u00f0A:14\u00de f24 \u00bc 32 E0pR2 Z b b Z a 0 xsRb b l sin5 hF1F2 \u00fems sin3 h cos2 hF2 III d ld b \u00f0A:15\u00de f25 \u00bc 16 E0pR2 Z b b Z a 0 xsR2 b l sin2 h sin 2hF2 2 \u00fems sin h cos h cos 2hF2 III d ld b\u00fe 4 E0pR2 Z b b Z a 0 j b l sin3 hF2 IId ld b \u00f0A:16\u00de f34 \u00bc 32 E0pR3 Z b b Z a 0 Rb b l sin5 hF1F2 \u00fems sin3 h cos2 hF2 III d ld b \u00f0A:17\u00de f35 \u00bc 16 E0pR3 Z b b Z a 0 R2 b l sin2 h sin 2hF2 2 \u00fems sin h cos h cos 2hF2 III d ld b \u00f0A:18\u00de f45 \u00bc 16 E0pR3 Z b b Z a 0 Rb b l sin3 h sin 2hF1F2 \u00fems sin2 h cos h cos 2hF2 III d ld b \u00f0A:19\u00de where the double integration limits are expressed as: a \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 b2 sin2 h q 1\u00fe r and b \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 \u00f01 r\u00de2 q = sin h" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001937_icems.2011.6073463-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001937_icems.2011.6073463-Figure6-1.png", + "caption": "Fig. 6. Schematic diagram of the 2-mass drive train modeling", + "texts": [ + " The equations of motion for 3-mass II model are given as 0 0 2 [ ( )] 2 [ ( )] [ ( )] 2 [ ( )] ( ) ( ) B B W BH BH BH B H B B H H BH BH BH B H LS LS LS H GBG H H GBG GBG LS LS LS H GBG GBG GBG G BH B H LS H GBG dH T K D D dt dH K D K D D dt dH K D D T dt d dt d dt \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u23a7 = \u2212 + \u2212 \u2212\u23aa \u23aa \u23aa = + \u2212 \u2212 + \u2212 \u2212\u23aa \u23aa\u23aa = + \u2212 \u2212 \u2212\u23a8 \u23aa \u23aa = \u2212\u23aa \u23aa \u23aa = \u2212\u23aa\u23a9 Where \u03c9GBG is the rotational speed of the mass stands for gearbox and generator; DGBG is the damping coefficients of the gearbox and generator. D. 2-Mass Drive Train Modeling Fig. 6 describes the 2-mass model, which is developed from 3-mass I model. Inertia HGBG is obtained by adding HGB and HG together. The equivalent shaft stiffness of the 2-mass system, KHLS, can be determined from the parallel shaft stiffness as in (6) [15]. Besides, the equivalent mutualdamping DHLS follows the same rules. The equations of motion for 2-mass model are given as 0 2 [ ( )] 2 [ ( )] ( ) BH BH W HLS HLS HLS BH GBG BH BH GBG GBG HLS HLS HLS BH GBG GBG GBG G HLS BH GBG dH T K D D dt dH K D D T dt d dt \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u23a7 = \u2212 + \u2212 \u2212\u23aa \u23aa \u23aa = + \u2212 \u2212 \u2212\u23a8 \u23aa \u23aa = \u2212\u23aa\u23a9 Where \u03b8HLS is the angle between the mass HBH and HGBG" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002153_20120215-3-at-3016.00208-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002153_20120215-3-at-3016.00208-Figure6-1.png", + "caption": "Fig 6. Hydraulic actuator", + "texts": [ + " Similarly, the mode at 85 Hz almost disappears in 11h (and also in 22h ), since the lower arms are very close to a node of the bending mode labelled \u20183\u2019. Figure 5 shows the W30 airframe, supported on its wheels with inflated tyres. The engines and main gearbox are mounted on a steel framework, which is itself supported by four elastomeric mounts on the roof of the cabin. An electrohydraulic actuator forms an integral part of the support system and imparts an internal force between the raft and the fuselage in parallel with each of the four elastomeric mounts as shown in Figure 6. The LMS system was used for modal tests. The transfer functions were measured and the eigenvalues and the modeshapes were determined. The measured transfer functions were curve-fitted as shown in Figure 7 for two of the measured receptances. The controller was designed to assign the poles of the system to the prescribed values using the theory of the receptance method. Real time implementation in dSPACE was carried out using gains fg determined from two different curve fitting methods. The details of the experimental work are given in the paper by Mottershead et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003958_icra.2014.6907468-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003958_icra.2014.6907468-Figure4-1.png", + "caption": "Figure 4. Reaction force matrix at the indentation depth of (a) 2 mm, (b) 4 mm and (c) 6 mm", + "texts": [ + " The result demonstrates the indenter was barely touching the reconstructed surface \u2013 hence, following the curvature of the tissue surface accurately during the entire scan process. We conclude that the tissue surface reconstruction can be used for indentation depth control during indentation scans that aim at acquiring a tissue\u2019s stiffness distribution. Three rolling indentation process were conducted with the indentation depths of 2 mm, 4 mm and 6 mm. During the process, the soft tissue surface is lubricated. Normal reaction force data was recorded (see Fig. 4). From the force matrices, one can see that the two nodules, A and B, are easily recognizable in the color-coded representation of the force matrix \u2013 the two nodules show as high force peaks (distinct red and yellow areas in an otherwise blue (low value) force distribution. IV. FEEDBACK TO THE USER A. Visualization of Tissue Deformation Deformation of the virtual soft tissue during palpation is displayed in real time using a geometrical deformable soft tissue model (see Fig. 5), which was established based on predefined finite element modeling considering the influence of the indenter diameter", + " In this experiment, haptic palpation could be seen as efficient as manual palpation. For this method to be employed in a real MIS setting, a smaller and sterilizable depth sensor or a binocular camera would need to be used instead of the Kinect. The two wrongly recognized nodule locations are both at the edges of the tissue model. The reason could be that the particular participant confused the changes of force caused by stiffness differences and the tissue texture. By just observing reaction force maps (like the one shown in Fig. 4), there is a risk of making mistakes in nodule identification and localization. In Fig. 4, there is an area with a relative high reaction force (see top right yellow area in Fig. 4 (a), (b) and (c)), which could be wrongly interpreted as a hard nodule if only the color coding of the shown force matrix was used in the analysis. In our human subject palpation experiments, users were able to detect the hard nodules correctly with the help of force feedback information. The reason for the slightly lower performance during the haptic palpation experiments compared to the manual palpation performance may be related to the limited tactile information experienced during the haptic feedback experiments" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000083_j.apm.2007.07.007-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000083_j.apm.2007.07.007-Figure2-1.png", + "caption": "Fig. 2. Proof of Eqs. (15) and (16).", + "texts": [ + " R1 df1 dr1 R2 df2 dr2 x\u03023 \u00bc 0: \u00f011\u00de These are the general equations that relate the positions of teeth 1, 2 to their shapes. Obviously, if the tooth shapes are known in the form of functions f1 and f2, then the path of contact and the TE, if any, can be uniquely calculated. Eq. (10) is solved in terms of R2f2 yielding: R2f2 \u00bc R1f1 a12: \u00f012\u00de Making use of a fundamental property of the translation matrix we obtain: R2f2k k \u00bc f2k k \u00bc r2 \u00f013\u00de thereby extracting from Eq. (12) the following calculation of r2: r2 \u00bc R1f1 a12k k \u00bc U 1\u00f0h1; r1\u00de: \u00f014\u00de It can be observed in Fig. 2 that vectors f2 and R2 f2 form by default an angle equal to h2, therefore f2 R2f2 \u00bc R2f2k k f2k k cos h2 ! h2 \u00bc cos 1 f2 R2f2 R2f2k k f2k k ! h2 \u00bc cos 1 1 r2 2f2 R2f2 : \u00f015\u00de An automatic solution process must be capable of selecting the correct sign for Eq. (15). This can be solved by observing that f2 R2f2 f2 R2f2k k x\u03023 \u00bc 1; \u00f016\u00de where x\u03023 is the unitary vector along the x3-axis (axis of revolution), so that Eq. (15) can be rewritten as: h2 \u00bc f2 R2f2 f2 R2f2k k x\u03023 cos 1 1 r2 2 f2 R2f2 \u00f017\u00de R2 is a function of h2, thereby rendering Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001143_j.commatsci.2009.11.013-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001143_j.commatsci.2009.11.013-Figure7-1.png", + "caption": "Fig. 7. Constraints of the model.", + "texts": [ + " The mild steel workpiece Q235 with dimensions 300 200 9 (mm3) are used for experiments and calculation. The lumped heat transfer coefficient acr , specific heat Cp, thermal conductivity k and rate of vaporization mer used in the model are temperature dependent. Their values are illustrated in Figs. 4 and 5. And other parameters used in computation are presented in Table 2. As shown in Fig. 6, Young\u2019s module and yield strength used in the model are temperature dependent. Poisson\u2019s ratio Nu is 0.33. And constrain of the workpiece is demonstrated in Fig. 7. First, thermal analysis is conducted to compute the temperature field in the workpiece for GMAW-P (test cases A1, A2) and Laser + GMAW-P hybrid welding (test cases B1, B2). Based on the temperature distribution, the weld geometry and size can be determined. Fig. 8 shows the computed and measured transverse crosssection of the Laser + GMAW-P hybrid weld (test case B1). It can be l stress distribution for test case B2 (Laser + GMAW-P). seen that hybrid weld is more laser-like at the bottom and more GMAW-like on the top due to the two process combinations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure9-1.png", + "caption": "Fig. 9. Kinematic joints for the rear suspension: (a) model with ideal joints; (b) model with bushing joints.", + "texts": [ + " The inertia properties of the rigid body, center of mass location and body fixed frame orientations, and all data for the suspensions, tires and stabilization bars, are described in the work by Verissimo [9]. Two models of the vehicle are considered in the study that follows: one with ideal kinematic joints in the suspen- sion systems and another using bushing joint in selected suspension elements. The location of the bushing joints on the rear suspension system, referred to as RB3, is represented in Fig. 9. The locations of the front suspension bushing joints are shown in Fig. 10(b) and c. The stiffness curves for the bushing joints model are obtained by a finite element model analysis, being the revolution bushing joints RB1, RB2 and RB3 the normal, tangential and rotational stiffness functions presented in Fig. 5. The application scenario, represented in Fig. 11, considers the vehicle riding over ten bumps, with a height of 0.1m. This case excites the roll motion of the vehicle chassis, making possible an evaluation of the suspension efficiency to reduce this chassis motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002405_s12046-012-0082-4-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002405_s12046-012-0082-4-Figure11-1.png", + "caption": "Figure 11. \u03b82 and \u03b8d2 with optimal PID gains.", + "texts": [], + "surrounding_texts": [ + "Consider a two-link manipulator as shown in figure 4 and its dynamics can be described by nonlinear equation (1). The matrices M(\u03b8), V (\u03b8, \u03b8\u0307) and G(\u03b8) for this two-link robot are M(\u03b8) = [ m11 m12 m12 m22 ] = [ a1 + a2 + 2a3 cos \u03b82 a2 + a3 cos \u03b82 a2 + a3 cos \u03b82 a2 ] , V (\u03b8, \u03b8\u0307) = [ \u2212 (a3 sin \u03b82) ( \u03b8\u03072 2 + 2\u03b8\u03071\u03b8\u03072 ) (a3 sin \u03b82) \u03b8\u03072 1 ] , G(\u03b8) = [ a4 cos \u03b81 + a5 cos (\u03b81 + \u03b82) a5 cos (\u03b81 + \u03b82) ] . (37) In the above expression a1, a2, . . ., a5 are constant parameters obtained from mass (m1, m2) and length (l1, l2) of robot links [ a1 = (m1 + m2) l2 1 , a2 = m2l2 2 , a3 = m2l1l2, a4 = (m1 + m2) l1g, a5 = m2l2g ] . The parameters are m1 = m2 = 1.0 kg, l1 = l2 = 1.0 m and g = 9.81 m/s2. For the system (10) with the expression of (37) we have the following numerical values. As1 = As2 = [ 0 1 0 0 ] , Bn1 = Bn2 = [ 0 5.8 ] , B1 = B2 = [ 0 1 ] , Cs1 = Cs2 = [ 1 0 ] , h(x) = [ h1(x) h2(x) ] = [ 1.72x2 12 + 1.68x2 22 + 3.36x12x22 \u2212 ( 5.12x2 12 + 1.72x2 22 + 3.44x12x22 ) ] , and x1 = [ x11 x12 ] = [ \u03b81 \u03b8\u03071 ] , x2 = [ x21 x22 ] = [ \u03b82 \u03b8\u03072 ] . (38) The stabilizing set of gains for links 1 and 2 for system (38), obtained by solving equations (19), (20) assuming values for Kdi from 1 to 100 and \u03c9 varying from 0.01 to 25 Hz are shown in figure 5. The shaded region in figure 5 is the stabilizing controller parameter space of joints 1 and 2. The set of controller gains for joints 1 and 2 are taken as K p1 \u2208 [ 10.1 500 ] , Ki1 \u2208 [ 10.1 500 ] , Kd1 \u2208 [ 10 100 ] , K p2 \u2208 [ 10.1 500 ] , Ki2 \u2208 [ 10.1 500 ] , Kd2 \u2208 [ 10 100 ] . (39) Knowing the ranges of controller gains (31) of the joints 1 and 2, genetic algorithm based optimization technique (Goldberg 1989) is used to maximize the fitness function J f given by J f = 1 1 + J , (40) where J = t\u222b 0 n\u2211 i=1 e2 i dt and ei (t) = \u03b8di (t) \u2212 \u03b8i (t). The optimal controller parameters are obtained for fixed as well as time-varying desired positions. The genetic operations used are arithmetic crossover, uniform mutation and ranking selection. The population size of 50 is taken and GA is run for 25 generations. 4.1 Case 1: Fixed desired positions Suppose the desired positions for joints 1 and 2 are \u03b8d1 = 30\u25e6 and \u03b8d2 = 45\u25e6. The combined optimal control parameters using genetic algorithm based optimizing technique for each joint are obtained as K \u2217 p1 = 176.83, K \u2217 i1 = 150.32, K \u2217 d1 = 86.84, K \u2217 p2 = 127.99, K \u2217 i2 = 128.4, K \u2217 d2 = 42.09. (41) The combined optimal control law for each joint is given by 4.2 Case 2: Time-varying desired positions Consider time-varying desired positions for joints 1 and 2 as \u03b8d1 = (1 \u2212 cos t) and \u03b8d2 = (1 \u2212 cos t). The combined optimal control parameters using genetic algorithm based optimizing technique for each joint are obtained as K \u2217 p1 = 129.11, K \u2217 i1 = 43.29, K \u2217 d1 = 56.31, K \u2217 p2 = 77.99, K \u2217 i2 = 20.4, K \u2217 d2 = 17.04. (43) The control laws with optimal controller parameters for both joints are obtained as Figures 6\u201313 reveal the effectiveness of the proposed decentralized PID control scheme and further it ensures tracking errors converge to zero asymptotically. 4.3 Stability analysis of two-link robot manipulator The stability analysis of the two-link robot manipulator (38) with the designed set of controllers (39) was studied by solving the LMI optimization problem (28) for all the corner matrices of Anew and Enew. The designed ranges of Anew and Enew are calculated using equations (25)\u2013(26) with the controller gains (39) and are given by Anew = diag {An1, An2} , Enew = diag {En1, En2} , (45) where An1 = \u23a1 \u23a3 0 1 0[ \u22122900 \u221258.6 ] 0 5.8[ \u2212500 \u221210.1 ] 0 0 \u23a4 \u23a6 , En1 = \u23a1 \u23a3 1 0 0[ 58 580 ] 1 0 0 0 1 \u23a4 \u23a6 , An2 = \u23a1 \u23a3 0 1 0[ \u22122900 \u221258.6 ] 0 5.8[ \u2212500 \u221210.1 ] 0 0 \u23a4 \u23a6 and En2 = \u23a1 \u23a3 1 0 0[ 58 580 ] 1 0 0 0 1 \u23a4 \u23a6 . As given in equation (27), the term wn(t , xw) can be bounded by a quadratic inequality and is constrained as wT n1(t, xw)wn1(t, xw) = ( 3.36x12x22 + 1.72x2 12 + 1.68x2 22 )2 \u2264 17.07x12x22 \u2264 8.54 ( x2 12 + x2 22 ) \u2264 xT w\u03b12 1 W T a1Wa1xw, wT n2(t, xw)wn2(t, xw) = ( \u22123.44x12x22 \u2212 5.12x2 12 \u2212 1.72x2 22 )2 \u2264 29.44x12x22 \u2264 14.72 ( x2 12 + x2 22 ) \u2264 xT w\u03b12 2 W T a2Wa2xw (46) (since x12 and x22 are much less than unity and the terms associated with the power of x12 and x22 equal to three or more than three are neglected), where \u03b11, \u03b12 > 0 and Wa1 and Wa2 are found out as Wa1 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0 0 0 0 0 0 2.92 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 2.92 0 0 0 0 0 0 0 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 , Wa2 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0 0 0 0 0 0 3.87 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 3.87 0 0 0 0 0 0 0 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 . (47) Two elements each of An1, An2 and Enew are of interval form, i.e., four corner matrices for each of An1, An2 and Enew are obtained. Table 1a shows the corner matrices of An1, An2 and Enew. Thus sixteen corner matrices are possible for Anew with the eight corner matrices of An1, An2. These corner matrices of Anew are given in table 1b. This sixteen combinations of Anew and four corner matrices of Enew are considered, thus there are sixty-four combinations for which optimization problem (28) is solved using LMI control toolbox (Gahinet et al 1995) with Wa1, Wa2 taken as equation (47). As the number of links increases the number of LMIs to be solved increases exponentially (23n where n is the number of links of the manipulator). Table 1c shows the values of \u03b11, \u03b12 obtained by solving the LMI problem (28) with the designed range of controller parameters given by (39). It is seen that a feasible solution exists for all the corner matrices. Hence, it is concluded that the set of decentralized PID controllers based on n1 = { } n2 = { } new = { } new = { } new = { } new = { } new = { } Kharitonov\u2019s theorem and stability boundary equation stabilizes the two-link manipulator system (38) with the numerical values of local controller parameters (39). The finite numerical values of \u03b1i , i = 1,2 indicate that the decentralized robust stability analysis of interconnected nonlinear system with its maximum nonlinear perturbations." + ] + }, + { + "image_filename": "designv11_3_0003525_tmag.2010.2089501-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003525_tmag.2010.2089501-Figure10-1.png", + "caption": "Fig. 10. Loss distribution obtained by proposed method.", + "texts": [ + " Therefore, the nonlinear time-stepping 3-D FEM is employed for the main calculation. Fig. 9 shows the meshes for the proposed method. The analysis region of the 3-D FEM is half the thickness of the axially segmented magnets [9] in order to take into account the magnet eddy currents, whereas the eddy currents in the laminated core are neglected. On the other hand, the number of 1-D elements is the same as that in the previous section. In this case, the computation time of the main 3-D FEM is 22 h 12 m, whereas that of the post 1-D FEM is 20 m. Fig. 10 shows the loss distribution. Fig. 11 shows the experimental and calculated iron losses including the magnet eddycurrent loss. The loss obtained by the 3-D FEM without the post 1-D FEM is also shown. From these results, the validity of the proposed method is confirmed. In this case, the 3-D FEM without the post 1-D FEM overestimates the loss, particularly when the rotational speed is low. This must have been caused by the overestimation of the high-frequency carrier-harmonic core losses, which form a considerable part of the total core loss, particularly at low-speed conditions [1]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000655_j.actaastro.2008.12.012-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000655_j.actaastro.2008.12.012-Figure1-1.png", + "caption": "Fig. 1. Coordinate axes of microsatellite in a circular orbit.", + "texts": [ + " In this paper, the microsatellite is assumed to be small and compact, without any flexible part. In addition to the body-fixed reference frame B, consider a local vertical local horizontal (LVLH) reference frame A with its origin at the center of mass of the microsatellite. The LVLH frame has a set of unit vectors { a 1 a 2 a 3}; with a 1, along the velocity direction of microsatellite in orbit plane, a 3, toward the Earth, and a 2, being the direction according to the right-handed Cartesian frame. The spatial directions of frames A and B are described in Fig. 1. To describe the orientation of the body-fixed reference frame B with respect to the LVLH reference frame A in terms of three Euler angles, the following successive rotations of Euler angles are applied: 1. Rotate the frame { a 1 a 2 a 3} about a 3 through the yaw angle to the frame { a \u2032 1 a \u2032 2 a \u2032 3}. 2. Rotate the frame { a \u2032 1 a \u2032 2 a \u2032 3} about a \u2032 2 through the pitch angle to the frame { a \u2032\u2032 1 a \u2032\u2032 2 a \u2032\u2032 3}. 3. Rotate the frame { a \u2032\u2032 1 a \u2032\u2032 2 a \u2032\u2032 3} about a \u2032\u2032 1 through the roll angle to the frame { b1 b2 b3}" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001315_978-3-540-30301-5_27-Figure26.13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001315_978-3-540-30301-5_27-Figure26.13-1.png", + "caption": "Fig. 26.13 (a) Multiple arm configurations that allow the manipulator to grasp the cylinder and (b) C-space visualization of the set of solution configurations", + "texts": [ + " This is one of the motivations and advantages of task space or operational space control techniques (Sect. 6.2). As an example, consider the task of grasping a cylindrical part with the planar three-DOF manipulator. Figure 26.11 shows two possible solution configurations yielding the same end-effector pose. The addition of obstacles to this workspace may cause one of the configurations to become disconnected from the zero configuration. These two configurations, however, may not be the only valid configurations that allow the manipulator to grasp the cylindrical part. Figure 26.13a shows several other configurations that allow the manipulator to grasp the part. By interpolating between these, an infinite number of different workspace poses that provide solutions to the grasping problem is generated. The inverse kinematics solutions associated with these workspace poses are contained in the continuous subset Cgoal \u2286 C, as shown in Fig. 26.13b, which represents the actual so- lution space to the planning problem for this reaching task. Adding the obstacle from Fig. 26.10 to the workspace of this manipulation task produces the configuration space depicted in Fig. 26.14. Note that the C-obstacle corresponding to collisions between the manipulator and the cylinder itself has been disregarded in this visualization for clarity. Approximately half of the possible solutions are disconnected from the zero configuration and a quarter are in collision, which leaves only one quarter as reachable and therefore suitable candidates for qgoal in a classical planning query" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure1-1.png", + "caption": "Fig. 1 Principle of operation", + "texts": [ + " The lubricating conditions at each moving interface in SWG were measured based on the contact electric resistance technique as a function of environmental pressure. In addition, a mixed lubrication analysis was conducted for simulating the lubricant film behaviour in the clearance between wave generator and flexspline. The SWG principle is unique in transmitting high torque through an elastically deformable component. JET415 \u00a9 IMechE 2008 Proc. IMechE Vol. 222 Part J: J. Engineering Tribology at IOWA STATE UNIV on October 13, 2014pij.sagepub.comDownloaded from The gear has only three concentric elements (Fig. 1). 1. The circular spline is a rigid ring with internal teeth, engaging the teeth of the flexspline across the major axis of the wave generator. 2. The flexspline is a non-rigid, thin cylindrical steel cup with external teeth on a slightly smaller pitch diameter than the circular spline. 3. The wave generator is a thin-raced ball bearing fitted onto an elliptical plug serving as a highefficiency torque converter. These three basic components function in the following manner. 1. The flexspline has a slightly smaller diameter and usually two fewer teeth than the circular spline" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001927_s12541-012-0021-7-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001927_s12541-012-0021-7-Figure4-1.png", + "caption": "Fig. 4 Screw involute surface of ZI worm (Left: Litvin and Fuentes)14", + "texts": [ + " In this study, gear geometry and Hertz\u2019s Law are employed to calculate the contact area between tooth surfaces. The worm and worm wheel have two principal curves respectively which are defined at a contact point as shown in Fig. 3, and this contact point is theoretically on the pitch circles of a worm and worm wheel. This study takes four principal curves into account to obtain the radius of curvature and uses Hertz\u2019s Law to calculate the contact area. Equations of four principal curves: Worm (Fig. 4, 5): A surface of a ZI worm is defined as14 cos cos sin sin cos cos sin tan b b b x r u y r u z u r \u03b8 \u03bb \u03b8 \u03b8 \u03bb \u03b8 \u03bb \u03b8 \u03bb \u2212 = \u2212 + = + = (4) sin cos cos cos cos sin tan b b b x r u y r u z r \u03b8 \u03b8 \u03b8 \u03b8 \u03bb \u03b8 \u03b8 \u03bb \u03b8 \u03bb = \u2212 + = + = (5) cos cos sin sin cos cos 0 b b x r u y r u z \u03b8\u03b8 \u03b8\u03b8 \u03b8\u03b8 \u03b8 \u03bb \u03b8 \u03b8 \u03bb \u03b8 = \u2212 \u2212 = \u2212 + = (6) where, b r = radius of the base cylinder, \u03bb = helix lead angle and u, \u03b8 = surface parameters. From the equations, the explicit form of a function of u is 1 cos sin cos b p p r r r u \u03bb \u2212 \u2212 = (7) where, rp = radius of the pitch circle of the worm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001855_j.compstruc.2008.12.014-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001855_j.compstruc.2008.12.014-Figure9-1.png", + "caption": "Fig. 9. Planar bending of a circular rod. The solid lines present the geometry of th", + "texts": [ + " As expected these values increase as the tip load is increased (and thus the deformations increase) or as the result of reducing the number of nodes. If the maximum elastic rotation component is smaller than 0.05, the error of the tip deflection calculation does not exceed 5%. In the case of the tip rotation, in order to obtain an error smaller than 0.1 it is necessary to have local elastic rotation components smaller than 0.02. As a first step of examining the capability of the new model to analyze large deformations of curved rods, the planar bending of a circular rod is studied as shown in Fig. 9. A circular rod is acted upon by a tensile force of magnitude 2F (when F is negative, compression is considered). Instead of solving the original problem, an equivalent problem of a rod that before deformation has the shape of a quarter circle, is considered (see Fig. 9). This quarter circle rod is clamped at the root and pulled by a force F at the tip. As the quarter ring deforms, the direction of the elastic axis at the free tip, where the force acts, is forced to coincide with the direction of the force (X direction). This is an indeterminate case that is solved by the procedure described in Appendix C. The tip coordinates, in the direction of the force and perpendicular to it, are la and lb, respectively (see Fig. 9). The problem was solved analytically by Bunce and Brown [44]. Fig. 10 presents a comparison between linear results, nonlinear results of the present transfer matrix model and analytical results from Bunce and Brown [44]. The normalized tip coordinates (dividing the dimensional coordinates by the rod radius R), are presented as functions of the normalized load, (F R2/E I). The linear model predicts only the X component of the displacement, while the normal displacement component is always equal to zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure9-1.png", + "caption": "Figure 9: Twin Cavity Supercharger Variable Drive", + "texts": [ + " Regarding the axial loads, if a second toroidal cavity is introduced, the two cavities react the loads against each other so removing the need for the thrust bearings. This increases the efficiency to ~89% (at high reaction loads) and correspondingly, increases the power capacity \u2013 for the same 50mm roller diameter, ~45kW power capacity results. The compromise is increased size and parts count. However, as the thrust bearings have been deleted, the impact on length of the variable drive system is minimised. The twin cavity design is shown in Figure 9 : The twin cavity design therefore provides a power dense variable drive in a package compatible with automotive supercharger applications. Traditional Turbo Compounding systems comprise a second power turbine located in the exhaust downstream of the primary turbine. This second turbine is mechanically connected to the crankshaft of the engine in order to recover energy from the exhaust that would have otherwise have been lost. Turbo compounding has been applied to airplane engines and in both on- & off-highway Commercial Vehicles (e" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001374_s12206-009-0101-5-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001374_s12206-009-0101-5-Figure1-1.png", + "caption": "Fig. 1. Schematic model of induction heating process.", + "texts": [ + " The architecture of the ANN model is optimized with a trial and error method, and the model is then trained by using both the results of the FEM analyses and the input heating parameters. The proposed network is tested and its results are compared with those of the experiments and the numerical analyses to reveal its feasibility to predict the deformation of steel plate in the forming process with induction heating. When high-frequency current flows through an inductor located over a conductor of a steel plate as shown in Fig. 1, a bundle of magnetic fluxes pass through the plate underneath, which is just to be coaxial coil, and consequently, eddy current is induced at the surface of the plate. The induced current generates electric-resistance heat in the plate, which can be used to bend the plate. The governing equation for the eddy current distribution in the induction heating process can be derived from Maxwell equations [10]. When the magnetic vector potential of the governing equation is analyzed, eddy current and heat-flux distribution can be calculated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001463_3.44295-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001463_3.44295-Figure7-1.png", + "caption": "Fig. 7 Moreland tire model.", + "texts": [ + " Its continued popularity is due in part to the ease with which it can be incorporated into the equations of motion of the landing gear system. Like the technique described in the body of this paper, the Moreland theory leads to a system of linear, constant coefficient, differential equations. Thus, the transfer functions are readily calculated in terms of ratios of polynomials in s. A comparison of these transfer functions with those derived from experimental data by Bode analysis provide the basis for assessing the validity of Moreland theory. The extended Moreland tire model can be represented by the mechanical system shown in Fig. 7. It consists of a rigid frame which carries a small, massless, rigid wheel. This wheel is attached to the frame by means of a linear spring ki and damper Ci. It is also restrained from rotating by the torsional spring kt> The frame in this model represents the actual tire wheel while the movement of the small rigid wheel corresponds to the motion of the center of the contact patch in the rolling tire. The relative displacement 2/i and rotation a\\ of this fictitibus wheel gives rise to the tire force and moment F = kiyi + Ciyi (Al) M = ktai (A2) In addition, the hypothesis is made that the cornering force is also related to the relative rotation a\\ according to the equation CyF = - (ai + (A3) In this equation, Cy is the tire yaw coefficient and C\\ is the tire time constant which accounts for the lag between F and ai. The negative sign indicates that a force F acting in the positive ?/i direction generates a negative or counter clockwise rotation ai. Finally one additional equation must be introduced which relates the wheel coordinates \\l/ and y to the displacement' y\\ and rotation a\\. This is obtained by assuming that the fictitious rigid wheel rolls without sliding in the ground plane, i.e., the instantaneous velocity vector V of the rigid wheel is always coincident with its plane. Thus, from Fig. 7, we have = Vy/Vx (A4) Introducing the assumption of small angles and expressing Vy in terms of y and yit one obtains = y (A5) In order to compare the transfer functions it is necessary to transform the independent variable in Eqs. (A1-A5) from time to the position coordinate x. Making this substitution and then taking the Laplace transform of Eqs. (Al), (A2), (A3) , and (A5) under zero initial conditions results in F(s) = (^ + C,7,\u00ab) M(s) = ktai(s) = -(I + dV + (A6) (A7) (A8) (A9) The cornering force response is obtained from Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001896_978-1-84996-432-6_87-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001896_978-1-84996-432-6_87-Figure5-1.png", + "caption": "Fig 5. CT test specimen", + "texts": [ + " Fixed dimensions are not given in these standards because all dimensions have to fulfill the special requirements of the linear-elastic fracture mechanics and with it they depend on material characteristics. For this reason only the interdependent dimensions are given. Not until all investigations are complete can the validation of the geometry be controlled according to special criteria and with it according to a valid critical stress intensity factor (KIc). All in all, six specimens of beam melted stainless steel (GP 1) are made with standard exposure strategies of the laser (cf. Figure 5 a). The chosen dimensions of the specimen are thickness B = 1 in \u00b1 0.02 in and width to load line W = 2 in \u00b1 0.01 in. Again the oversized specimens are machined after beam melting to the desired geometry and prepared for testing (cf. Figure 5 b). The fracture toughness test begins with the initiation of the flaw by cyclic loadings (10 Hz) of 20,000 N and 15,000 N with stress ratios of R = 0.1. The loading is transferred to the specimen by bolts in the boreholes. The fatigue crack initiates and propagates starting at the top of the Chevron-notch and is stopped after 4-5 mm. The crack propagation is controlled by measuring the voltage potential at the two electrodes on the left side of the specimen in Fig. 5 b. Before carrying out the tensile test, the two electrodes are replaced by a displacement transducer for measuring the crack opening. According to DIN EN ISO 12737, the tensile test is carried out at a rate of rise of the critical stress intensity factor of 0.5\u20133.0 MPa\u221am/s. After the specimen is broken, several dimensions such as thickness (B), width to load line (W) and the mean crack length (a) are measured from the photographs as illustrated in Figure 6. The crack surface of the fatigue fracture is smooth and light whereas the surface of the final fracture is rough and dark" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003071_j.reactfunctpolym.2012.10.008-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003071_j.reactfunctpolym.2012.10.008-Figure6-1.png", + "caption": "Fig. 6. Two types of geometry for T-NE ribbon specimens.", + "texts": [ + " This result indicates that the thermally induced bending behavior of H-NEs is quantitatively pre- dictable on the basis of the thermal deformation data of the corresponding NEs with planar or vertical alignment. A finite element approach to consider more rigorously the strain distribution in the thickness direction is described elsewhere [11]. Thermal deformation of the NEs with twist director configurations (T-NEs) was investigated for the ribbon specimens with the two types of geometry designated as L- and S-geometries (Fig. 6). The director at the plane in the middle of the two surfaces in the L- and S-geometries is parallel to the long or short axis of the ribbon specimens, respectively. Further, the dimensional effect on the shape formation was examined using the ribbon specimens with various widths (ca. 0.2\u20130.8 mm) but with constant thickness (35.2 lm) in the flat state. We observe that the T-NE ribbons select a helicoid or spiral ribbon depending on the width [12]: The narrow ribbon forms a helicoid with Gaussian suddle-like curvature (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000691_s0025654408030059-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000691_s0025654408030059-Figure1-1.png", + "caption": "Fig. 1.", + "texts": [ + " On this plane, the stationary motions are associated with the set \u03a3 = \u03a3+ \u222a \u03a3\u2212 \u222a \u03a30, where the \u03a3\u00b1 are the straight lines q = \u22131 + p2 2(1 \u2213 b)2 (q = f(\u00b11)), (16) the set \u03a30 consists of curves determined parametrically by the formulas (see (14), (15)) q = f(x, p2), p2 = \u03d5(x), (17) and x \u2208 (\u22121, 1) is a parameter that should be eliminated from relations (17). The shape of the Smale diagrams significantly depends on the top parameters a and b. On the basis of a detailed analysis of the function \u03d5(x), the plane of these parameters is divided into seven domains (a)\u2013(g) (Fig. 1) each of which is associated with its own Smale diagram (Fig. 2 a\u2013g). Domains (a)\u2013(g) are determined by the relations (a) a > 1 + b; (b) 1 + b > a > \u03b1+(b); (c) \u03b1+(b) > a > \u03b1\u2217(b); (d) \u03b1\u2217(b) > a > \u23a7 \u23a8 \u23a9 1 \u2212 b, b \u2265 7 \u2212 \u221a 48, \u03b1+(\u2212b), b < 7 \u2212 \u221a 48; MECHANICS OF SOLIDS Vol. 43 No. 3 2008 (e) \u03b1+(\u2212b) > a > 1 \u2212 b; (f) 1 \u2212 b > a > \u03b1\u2212(\u2212b); (g) a < \u23a7 \u23a8 \u23a9 1 \u2212 b, b \u2265 7 \u2212 \u221a 48, \u03b1\u2212(\u2212b), b < 7 \u2212 \u221a 48. Here \u03b1\u00b1(b) = 1 \u2212 b 8 (7 + b \u00b1 \u221a 1 + 14b + b2), \u03b1\u2217(b) = (1 + 3b2)(1 \u2212 b2) 1 + 6b2 + b4 . On the Smale diagrams (Fig", + " We note that all points belonging to the set \u03a3 on the plane (p2, q) are invariant with respect to the MECHANICS OF SOLIDS Vol. 43 No. 3 2008 phase flow of system (1)\u2013(4) for M = 0, and all other points evolve along the line p2 = const in the direction of decreasing q. This fact allows us to perform global qualitative analysis of the dynamics of a dynamically symmetric ball with displaced center of mass on a plane with sliding friction. For example, let the top parameters belong to the domain (b) (Fig. 1); in this case, the generalized Smale diagram has the form shown in Fig. 2 b. If we set the top in a position close to the stable equality \u03b3 = e and let it spin fast (p2 = p2 0 > p2 \u2212 > p2 +) about its symmetry axis, then on the plane (p2, q) such initial conditions are associated with a point lying in a neighborhood of the line \u03a3+ to the right of the line p2 = p2 \u2212. If this point lies below (above) the line \u03a3+, then its starts to move along the line p2 = p2 0 in the direction of decreasing q and uniquely (with probability 1) tends to the point p2 = p2 0 lying on the line \u03a3\u2212; i", + " Finally, if we set the top in a position close to stable equilibrium and let it spin about the symmetry axis at a small (p2 = p2 0 < p2 +) angular velocity, then it continues stable rotations about the vertical symmetry axis for the lowest position of its center of mass. We note that, with probability 1, these final motions of the top depend only on the initial value of the constant of the Jellett integral (the parameter p2) and are independent of the initial value of the total mechanical energy (the parameter q). Thus, for almost all initial values, the final motions of the top whose parameters lie in domain (b) (Fig. 1) are uniform rotations about the vertical symmetry axis for the lowest (if p2 0 < p2 +) or the highest (if p2 0 > p2 \u2212) position of the center of mass or precession (if p2 0 \u2208 (p2 +, p2 \u2212)) motions. But if the top parameters lie in the domains (c)\u2013(e) (Fig. 1), then, just as in the above case, the fast rotation of the top about the vertical symmetry axis is stable (unstable) for the highest (lowest) position of the center of mass (fig. 2 c\u2013e). But, in contrast to the case considered above, one can predict the final motions of the top whose parameters lie in domains (c)\u2013(e) with probability 1 only for p2 0 < p2 \u2217 or p2 0 > p2 + (in cases (c) and (d)) and for p2 0 < p2 \u2212 of p2 0 > p2 + (in case (e)). If these inequalities are violated, it is impossible to predict the final motions of the top whose parameters lie in domains (c)\u2013(e) with probability 1, because for the same value p2 0 \u2208 (p2 \u2217, p 2 +) (in cases (c) and (d)) and p2 0 \u2208 (p2 \u2212, p2 +) (in case (e)), there exist two stable steady motions of the top", + " 3 2008 Conditions (20) mean that the total mechanical energy of the top decreases on motions without sliding much more slowly than on motions with sliding, and the modulus of the projection of the angular momentum onto the position vector of the point of contact decreases slowly for all motions. Obviously, the friction torque M destroys the steady motions (12), but under conditions (20), system (1)\u2013(4) admits quasisteady motions close to steady ones. Therefore, the above qualitative analysis of the top dynamics without the friction torque taken into account can be considered as the generating case of analysis. Consider the top whose parameters lie in domain (b) (Fig. 1). We set the top in stable equilibrium or in a state close to it and let it spin fast about the symmetry axis (p2 0 > p2 \u2212 + \u03b4, \u03b4 > 0). Such initial conditions are associated with point Q on the plane (p2, q) (Fig. 2 b) lying in a neighborhood of the line \u03a3+ on the vertical p2 = p2 0. Under the action of the phase flow of the system (1)\u2013(4), this point starts to move in the direction of decreasing q (\u201cfast\u201d) and p2 (\u201cslow\u201d) until it comes to the line \u03a3\u2212 to the right of the line p2 = p2 \u2212 (if \u03b4 > 0 is sufficiently large)", + " After this, the top starts precession motions; in this case, the angle of inclination of its symmetry axis from the vertical starts to increase \u201cslowly\u201d from zero (for p2 = p2 \u2212) to 180\u25e6 (for p2 = p2 +). Then the top starts to rotate \u201cslowly\u201d in the original position so that its angular velocity \u201cslowly\u201d decreases to zero; finally it stops in stable equilibrium. The above qualitative analysis of the top dynamics agrees with both numerical [7] and natural experiments. Similarly, one can (qualitatively) study the influence of the friction torque on the dynamics of a top whose parameters lie in domains (a)\u2013(e) and (a), (f), and (g) (Fig. 1). To obtain quantitative estimates, it is insufficient to have conditions (5) and (19) on the force F and the torque M. To this end, it is required to have their explicit expressions given in [3]. The research was financially supported by the Russian Foundation for Basic Research (projects nos. 07-01-00290 and 06-08-01574). 1. P. Contensou, \u201cCouplage Entre Frottenment de Glissement et Frottenment de Pivotement Dans la The\u0301orie de la Toupie,\u201d in Kreiselprobleme Gyrodynamics (Springer-Verlag, Berlin, 1963; Mir, Moscow, 1967), pp" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003151_ecce.2013.6647071-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003151_ecce.2013.6647071-Figure1-1.png", + "caption": "Figure 1. Winding configuration with neutral point (NP connected).", + "texts": [ + " The current is measured in six coils (U1, U2, V1, V2, W1, and W2) directly because a circulating current occurs only in a closed circuit and cannot appear in the input terminal of three phases. Therefore, the prototype motor developed for this paper has the structure in which the stator coil can connect freely and that has 12 output terminals including U1+ and U1\u2013. Six current probes are used to measure the current in the coils, i.e., U1+, U2+, V1+, V2+, W1+, and W2+ at the same time. These outputs are then saved in a data logger. The operating conditions are under no load operation at a frequency of 50 Hz. Figs. 1 and 2 are diagrams of stator winding equivalent circuits: Fig. 1 shows the circuit with the connected neutral point (NP connected) while Fig. 2 shows that with the unconnected (NP non-connected). Both types use the same motor, which can be switched to neutral point on or off. Fig. 3 outlines the static rotor eccentricity where the rotor is displaced vertically toward the U2 coil by 13 % of the normal air-gap length. In addition, the experimental IPM has a maximum erection tolerance of 7 % and hence the total eccentricity value ranges from 13 up to 20 %. III. VOLTAGE DIFFERENCE UNDER ECCENTRICITY This paper introduces a theoretical expression of the circulating current for the purpose of comparing connected and unconnected neutral points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002115_13552541211193485-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002115_13552541211193485-Figure4-1.png", + "caption": "Figure 4 Measurement of flatness using reverse engineering", + "texts": [ + " The flatness and cylindricity calculations are based on the ASME Y14.5M-1994 dimensioning and tolerancing standard. Cylindricity is defined as the distance of separation between two concentric cylinders within which all the irregularities of the cylindrical surface should lie. Hence a value of 0mm for cylindricity represents a perfect cylinder. Flatness is defined as the distance of separation of two planes within which all the irregularities of the surface should lie. Cylindricity and flatness are measured using a white light scanner which is shown in Figure 4(a). During measurement, white-light stripes are projected over the object to be measured. Furthermore, to increase accuracy, they are emitted onto the measured object in a selected stripe code (gray code) and in phase shift. These white-light stripes are reflected back from the surface of the object and are recorded by two charge-coupled device (CCD) cameras. Thus, after triangulation, the 3D coordinates can be calculated for every point on the surface. After the measurement, raw data are obtained in the form of a 3D scatter plot. The point cloud data obtained after scanning (Figure 4(b)) is post processed using Polyworkse software. A cylinder is fitted and cylindricity values are calculated on the cylindrical surface of the fitted cylinder using Polyworkse GD&T module. Also flatness values are calculated from the plane fitted to the points falling on the flat face of the cylinder. The measured values of cylindricity and flatness have been tabulated for each run in Table III. Deviation plots have been also obtained by comparing the nominal CAD data and measured data (Figure 4(c))[2]. The accuracy of the GD&T measurements Modeling and minimization of form error in SLS prototyping K. Senthilkumaran, P.M. Pandey and P.V.M. Rao Rapid Prototyping Journal Volume 18 \u00b7 Number 1 \u00b7 2012 \u00b7 38\u201348 using white-light scanner and Polyworks software together is ^13 mm[3]. 3.5 Development of models for flatness and cylindricity The objective of the present work is to study the influence of process and geometry parameters on geometric errors and to build a model relating these errors with the process parameters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000076_j.robot.2008.01.002-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000076_j.robot.2008.01.002-Figure3-1.png", + "caption": "Fig. 3. Locked joint failure: (a) lateral view of the locked failure at joint 1, (b) lateral view of the locked failure at joint 2 (or joint 3) and (c) front view of the locked failure at joint 2 (or joint 3).", + "texts": [ + " Hence a locked joint failure happening to a leg with three degrees of freedom will result in twodimensional motions of the leg and thus will forbid the failed leg to have normal swing in the transfer phase or backward movement in the support phase. Nevertheless, unlike free-swinging failure [12] and mutilation failure, a locked joint failure does not take away body-supporting ability from the failed leg. For employing the failed leg in post-failure walking, we should examine the configuration of the failed leg determined by the position of a locked joint and the resulting change of the reachable area. Fig. 3 illustrates the behavior of a failed leg with the geometry of Fig. 2. After joint one of a leg is locked from failure, the kinematics of the failed leg is the same as a two-link revolute joint manipulator. Its workspace is reduced to the plane made of the two links and the reachable region of the foothold position in the working area is projected onto a straight-line of which the lateral view is shown in Fig. 3(a). \u03b8\u03021 denotes the locked angle of joint one, and the values of \u03b82 and \u03b83 are determined by the foothold position. The locked failure at joint two or joint three yields almost identical post-failure behavior. When joint two or three is locked, the failed leg is tantamount to a one-link manipulator with two revolute joints. Since the altitude of the robot body is assumed to be constant, if one of joints two and three is locked, the angle of another joint is fixed too for a given foothold position. For instance, joint two is assumed to be locked at \u03b8\u03022 and the failed leg is placed onto the point P in Fig. 3(b). Though the leg might have another foothold position P \u2032 in kinematics, it is impossible to place onto the point because the altitude of the robot body does not change. Therefore, joint three is also \u201clocked\u201d at \u03b8\u03023 in the configuration of Fig. 3(b), and the failed leg moves by swinging joint one as shown in Fig. 3(c), making the track of foothold positions in the shape of an arc. When the leg trajectory is a straight-line, the above characteristics of locked joint failures makes the failed leg have only one possible foothold position. Hence, in our study, a given configuration of the locked joint failure can be represented by a point on the workspace. We define such a point as the coordinate on the axis that has the direction of the leg trajectory and the origin at the center point Ci . Fig. 4 shows our definition, where ri is the position of a failed leg on the designated axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001895_978-3-642-32448-2_11-Figure11.2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001895_978-3-642-32448-2_11-Figure11.2-1.png", + "caption": "Fig. 11.2 Modeling of the robot with tilting at virtual axes", + "texts": [ + " wrep i i i i i i i link Trans p Rot z Trans d p Trans a Rot x \u03b8 \u03b1 = \u2212 This model is extended by two variable rotations around axes, orthogonal to the joint's axis, leading to the final extended kinematics model , , ,(0,0, )\u00b7 ( ; )\u00b7 ( ; )\u00b7 ( ; )\u00b7 (0,0, )\u00b7 \u00b7 (0,0, )\u00b7 ( ; ) ext i i i x i y i i i i i link Trans p Rot z Rot x Rot y Trans d p Trans a Rot x \u03b8 \u03b8 \u03b8 \u03b1 = \u2212 with \u03b8x,i and \u03b8y,i being the rotation angles caused by elasticities around the additional axes. These additional revolute axes are referred to as virtual joints (cf. Fig. 11.2). It should be noted that these virtual joints are optional. One or both virtual joints can be added to each link, if additional elasticities need to be modeled for this link. Additional Dynamics Properties Simulating the motion dynamics of the robot requires additional parameters for each link of the robot. Namely, these are the mass mi , the inertia tensor Ii and the center of mass comi for each link. In the implementation used for this work, the center of mass is described with respect to the coordinate frame defined by linkext,i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002524_3.5535-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002524_3.5535-Figure5-1.png", + "caption": "Fig. 5 Solution curves in neighborhood of origin;", + "texts": [ + " Figure 4 summarizes, schematically, the character of the four singularities summarized in Table 1 plus the higher-order singularity at the origin deduced from limiting cases of Eq. (24). The relative spacing between the singularities is grossly distorted, since e is, in general, a very small quantity compared with a. The qualitative behavior of the solution curves in the neighborhood of the singularities near the origin, as determined from a plot of Eq. (24) by the method of isoclines, is shown in Fig. 5. The nature of the crossover from an oscillation about the positive branch of the pitch frequency curve to an oscillation about the negative branch now becomes apparent. Figure 5 shows that there are two switching lines in the positive halfplane and one in the negative halfplane. To the right of the first switching line, an oscillation about the positive branch persists until the amplitude reaches sufficient magnitude to reach the first switching line as y -> 0. It then appears possible to follow a trajectory between the two switching lines that will pass through the singularities but continue around the positive branch (trajectories labeled 3 and 4). Beyond the second switching line the trajectory loops around the negative branch (y = \u2014a) and can return to the positive branch provided it passes to the right of the switching line in the negative halfplane (trajectories 1 and 2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000799_00221287-115-1-59-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000799_00221287-115-1-59-Figure4-1.png", + "caption": "Fig. 4. Effect of L-methionine-DL-sulphoximine on the accumulation of arginine (0), glutamine (A) and glutamic acid (G). Conidia of the po l -3 strain were incubated in MM without NH4N08 but supplemented with glutamine (100 pg ml-l) and methionine sulphoximine (5 m), after 12 h preincubation in nitrogen-free MM.", + "texts": [ + " When conidia of theprol-3 strain were incubated in a nitrogen-free medium in the presence of MS, this compound was degraded and glutamine and arginine accumulated (Table 2). This accumulation was observed only after 12 h incubation and no glutamine synthetase activity was detected during this lag phase. It is probable that this period was required for MS degradation. Similar effects were observed in the presence of MS plus NH,NO,, but glutamine and arginine accumulated on addition of glutamine plus MS to nitrogen-deprived prol-3 cultures (Fig. 4). 5 M I C 115 64 G. ESPfN, R. PALACIOS A N D J. MORA As shown above, double mutants lacking glutamine synthetase activity are unable to utilize the nitrogen of glutamine to synthesize and accumulate arginine (Fig. 2). The conversion of glutamine into arginine under conditions in which the activity of glutamine synthetase is completely inhibited (Fig. 4) indicates that MS is being degraded. Because of its competition as a substrate for the transaminase, MS probably spares glutamine for arginine synthesis. The degradation of MS could provide the glutamic acid required for arginine synthesis. Eflect of nitrogen deprivation and the gln-I mutation on the catabolism of arginine We have previously reported that glutamine prevents the catabolism of arginine by repressing arginase synthesis ; thus proline auxotrophs will not grow on arginine if glutamine is simultaneously present in the culture medium (Vaca & Mora, 1977)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002505_12.918523-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002505_12.918523-Figure4-1.png", + "caption": "Fig. 4 Example of 2-DOF space manipulator approaching a square target for capturing", + "texts": [ + "aspx , x \u03bb H H x \u03bb (14) with the initial and final conditions given in (12). For our problem, the final time and final motion state of the optimal control are known. In other words, it is a fixed time and fixed boundary problem. However, because of its complex nonlinear nature, it still has to be solved numerically. To show the application of the afore-mentioned optimal control strategy, we present an example using a 2-DOF planar manipulator in this section. The parameters of the satellites and robot are defined in Fig. 4 and Table 1. The system is assumed as shown in Fig. 4, the center of mass of the servicing system including the robot can be found to be: 0 0.2 0.41 0 m n i i Ti m C m r We also assume that Proc. of SPIE Vol. 8385 83850J-8 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 11/02/2015 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx 20 0.5 0 m s T a 1 0 0 (0) 0 1 0 , (0) [2.117 1.41 0] m, 0 0 1 =0.147 rad/sec, (0) 0.735 0 0 m/s T T t R r \u03c9 v In the 2-D case, the rotational motion is simple and thus, we can easily find that the optimal time and orientation of the target satellite for the robot to capture with zero attitude impact are 4ft s ( ) ( )f tt v R \u03c9 a [ 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000687_j.triboint.2008.09.005-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000687_j.triboint.2008.09.005-Figure2-1.png", + "caption": "Fig. 2. Structure of test rig: (a) outline photos and (b) schematic diagram.", + "texts": [ + " In addition, MoS2 is uniformly coated on the surface of top foil before experiment. As a kind of lubricant, MoS2 has very small friction coefficient and good lubrication performance at working conditions. Previous experiments for CFB showed that coating MoS2 can obviously reduce wear of the friction pair used in this paper [8]. As shown in Fig. 1, one end of foil element is fixed on tile, and another end is free. The surface of tile is preformed into an inclined plane followed by a plat plane, which can form a convergent structure with the thrust face. Fig. 2 shows the structure of test rig, which includes two sections, rotation part and loading part. To minimize the possible unstable effect caused by journal bearing, aerostatic gas bearings are chosen in rotation part. The rotor is of length 177.5 mm with a shaft in diameter of 17 mm, and driven by an impeller with diameter of 25 mm. The weight of rotor is about 350 g. Flow rate of turbine is controlled by two valves for coarse adjustment and fine adjustment, respectively. Test bearing is fixed on a piston connecting a rod which can apply axial load on the bearing by compressed gas" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.30-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.30-1.png", + "caption": "Fig. 2.30 The two levels of pallet fork in the English lever escapement", + "texts": [ + " Mudge was apprenticed to George Graham. It is one of the earliest escapements that does not require a pendulum and hence, is regarded as a milestone in the history of watch and clock. As shown in Fig. 2.29, the English lever escapement is composed of four parts: the escape wheel, the anchor-like pallet fork, the balance wheel and the hairspring. The axes of the balance wheel, the pallet fork and the escape wheel form a rightangled triangle (as shown by the dot-dash line). Note that the pallet fork has two levels as shown in Fig. 2.30: Level 1 is the balance wheel level, on which the balance wheel has a half-cycle shaped ruby, called the impulse pin, to turn the pallet fork. Level 2 is the escape wheel level, on which the escape wheel turns the two rubies on the pallet fork, called the entry pallet jewel and the exit pallet jewel. There are also two position pins that limit the swing of the pallet fork. The big innovation of the English lever escapement is the use of the hairspring. It allows the balance wheel swinging in a large angle and hence, is much more reliable" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000779_bfb0109667-Figure2.2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000779_bfb0109667-Figure2.2-1.png", + "caption": "Fig. 2.2. Planning a connecting contour. For a trajectory to be feasible, it must pass through the origin of the moving plane (current position of the vehicle) and it must be oriented along the x axis. A connecting contour is an approximation to the target contour that simultaneously satisfies the feasibility constraints.", + "texts": [ + " Needless to say, we cannot count on the vehicle ever being exact ly on the contour. However, causali ty and the non-holonomic constraint imposes tha t any feasible t ra jec tory must go through the current position of the vehicle, and it must have its tangent oriented along its x-axis. Since the target contour is known only locally through the measurements V ( x i , t ) , i = 1 . . . N , one could imagine an \"approximation\" to the target contour which, in addit ion to fitting the measurements ~'(xi , t) , also satisfies the two addit ional nonholonomic constraints (see figure 2.2). We call such approximat ing t ra jec to ry a \"connecting\" contour 1. For the case of a wheel, the connecting contour would s tar t at the current position of the vehicle with the tangent pointing along the x -d i rec t ion , and end at a point (Xc, V(xc, 0)) on the contour with the same tangent. Overall the connecting contour Vx must satisfy the minimal set of conditions: c(0) = 0 = 0 (2 .5 ) c(xc) = (Xc,0) = The simplest curve tha t satisfies the four above conditions is a polynomial of degree 3", + " By construction, the vehicle is 1 The choice of the connecting contour depends upon the differentially flat structure of the system. A connecting contour for a flat system of order p (i.e. the flat outputs need to be differentiated p times to recover the state) must satisfy p causality conditions and be of class C p- 1. For example, the connecting contour for a vehicle with M trailers must satisfy at least M + 2 causality conditions. on such a contour, and oriented along its tangent. Therefore, one may hope to be able to apply the exact tracking controller (2.4), where the curvature is that of the connecting contour (see figure 2.2): w(t) - v ~ (0, t) (2.6) This strategy is bound to failure for several reasons. First the composite contour is not a feasible path for the vehicle since continuity of the secondderivative (and therefore of the control) is not guaranteed at xc. Second, while the connecting contour is being planned, the vehicle may have moved, so that the initial conditions (2.5) are violated. More in general, the controller should be updated in response to added knowledge about the contour whenever a new measurement comes in, aking to a feedback control action" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001660_j.corsci.2009.02.004-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001660_j.corsci.2009.02.004-Figure7-1.png", + "caption": "Fig. 7. SEM micrographs showing microstructure at various locations of LBW weldment following corrosion test: (a) FZ, (b) FZ at higher magnification, (c) WDZ, and (d) WDZ at higher magnification.", + "texts": [ + " This phenomenon is commonly referred to as \u2018\u2018weld decay\u201d and is both a common and a serious problem in arc-welded austenitic stainless steels [10\u201312,16]. As shown in Fig. 6(d), a similar effect is observed in the HAZ of the current Alloy 690 GTAW weldments. Compared to the CGZ, it can be seen that the WDZ contains deep grooves along the grain boundaries and shows clear signs of IGC. Finally, Fig. 6(e) shows that the BM of the GTAW weldment is not affected by IGC since it experiences a relatively lower peak temperature during the welding process. Fig. 7(a) and (b) present SEM images of the FZ in the convex side view of the LBW U-bend specimen. Comparing these images with those presented in Fig. 6(a) and (b) for the equivalent region of the GTAW specimen, it is evident that the FZ of the LBW weldment has a far better IGC resistance. This phenomenon reflects the fact that the FZ in the LBW specimen has a significantly lower heat input, a more rapid cooling rate, and is fabricated without using a filler metal. The impurities enriched at grain boundaries during solidification of weld metal (i", + " In the study by Lu et al. [34], a low heat input and high cooling rate of LBW is likely to reduce the microsegregation of alloying elements and the formation of Cr-depleted zones, resulting in improvement of the resistance to localized corrosion. Other related references also showed the same effect of the LBW process [16,23]. Thus, the fine microstructure of the FZ produced by LBW process is less susceptible to localized corrosion, as compared to those coarse micro- structures produced by the GTAW process. Fig. 7(b) shows the tiny pitting holes that formed at the grain boundaries. Fig. 7(c) and (d) present SEM images of the WDZ of the convex side view of the LBW U-Bend specimen. It is observed that this region of the weldment also shows no obvious signs of IGC. In earlier studies, Mayo [35] and Sahlaoui et al. [36] examined the Cr depletion profiles and IGC/IGSCC susceptibility of Alloys 600 and 690 at different aging temperatures and times. But the effects of different aging temperatures and times on the time\u2013temperature transformation kinetics of Cr carbides and references related to Alloys 600/690 have seldom been reported" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000331_20070829-3-ru-4911.00048-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000331_20070829-3-ru-4911.00048-Figure9-1.png", + "caption": "Fig. 9. Diagram of the aircraft body axes system.", + "texts": [ + " We consider the problem of controlling a rigidbody aircraft so that it follows a predefined path in space with desired speed and attitude. The control actuation is assumed to be represented by three surface deflection angles (corresponding to ailerons, elevators and rudder) and the thrust. The variables that are assumed to be measurable are position, velocity, Euler angles and angular rates. The mathematical model used in the design is that of a six-degrees-of-freedom rigid body of constant mass, whose attitude is expressed in standard Euler angles (roll, pitch and yaw) based on a conventional body axes system, depicted in Figure 9. The proposed control scheme consists of the following modules. 1. An adaptive attitude controller that achieves asymptotic tracking of the roll, pitch and yaw angles, in the presence of aerodynamic moments with unknown coefficients. 2. An adaptive airspeed controller that achieves asymptotic regulation of the airspeed, in the presence of unknown drag and lift coefficients. 3. A trajectory tracking controller that produces the required references for the roll, pitch and yaw, in order to asymptotically follow a predefined geometric path" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000945_978-1-4020-8600-7_5-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000945_978-1-4020-8600-7_5-Figure3-1.png", + "caption": "Fig. 3 The 5-dof parallel manipulator.", + "texts": [ + " (16), (17) and (22) that the condition for constraint singularity involves all joint axes, which fact implies that the mechanism remains in constraint singular configuration even when the driven joint is changed. In this section a type of parallel manipulator with five degrees of freedom is introduced and its singular configurations are analyzed using the geometric algebra. The considered parallel manipulator has four legs; the first leg has RRPRR structure and the other three legs have identical SPS structure (Figure 3). The driven (active) joints are the four prismatic joints of the legs and a revolute joint R attached to the base of the RRPRR leg. In this case, the RRPRR (R \u22a5 R \u22a5 P \u22a5 R \u22a5 R) leg has two driven (active) joints: the first one (R \u2013 revolute joint attached to the base) and the prismatic joint (P). The SPS legs have full mobility and each one has one driven joint (the P joint). In this case the SPS (or UPS) leg has only one possible dual screw, or reciprocal screw to the joint axis associated with the U and S joints", + " The RRPRR leg has five degrees of freedom and in order to have full mobility one extra dummy joint (denoted by a superscript d in the equations) is added, which can be considered as active but locked. The dual screws associated with the active and dummy joints of the first (RRPRR) leg are as follows: 1D1 = (1S2 \u2227 1S3 \u2227 1S4 \u2227 1S5 \u2227 1Sd)I \u22121 6 , (23) 1D3 = (1S1 \u2227 1S2 \u2227 1S4 \u2227 1S5 \u2227 1Sd)I \u22121 6 , (24) 1Dd = (1S1 \u2227 1S2 \u2227 1S3 \u2227 1S4 \u2227 1S5)I \u22121 6 , (25) 45 3.2 Singularity of a 5-dof Parallel Manipulator 46 T.K. Tanev where 1S1 is the axis of the joint attached to the base and 1S5 is the axis of the joint connected to the moving platform of the RRPRR leg (Figure 3). Applying the identities of the geometric algebra for the outer product of the duals from Eqs. (23), (24) and (25) one obtains: D13d = 1D1 \u2227 1D3 \u2227 1Dd = \u03bb(1S2 \u2227 1S4 \u2227 1S5)I \u22121 6 , (26) where \u03bb = (1S1 \u00b7 1D1)( 1Sd \u00b7 1Dd) is a scalar; the above result is obtained bearing in mind that 1Di \u00b7 1Sk = 0 (i = k) and 1Di \u00b7 1Sk = 0 (i = k). In fact, the 3-blade from Eq. (26) is a blade of non-freedom for the RRPRR leg. One of the wrenches of constraints (1Cd = 1D\u0303d ) is uniquely defined by Eq. (25). The algebraic condition for singularity can be written as follows:{[(1S2 \u2227 1S4 \u2227 1S5)I \u22121 6 ] \u2227 2D \u2227 3D \u2227 4D } I\u22121 6 = 0, (27) where 2D = 2C\u0303, 3D = 3C\u0303 and 4D = 4C\u0303 are duals associated with the three SPS legs: jD = j S1 \u2227 j S2 \u2227 j S4 \u2227 j S5 \u2227 j S6 (j = 2, 3, 4)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001797_aqtr.2010.5520862-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001797_aqtr.2010.5520862-Figure4-1.png", + "caption": "Figure 4. The sensitivity of IR sensors.", + "texts": [ + " The sensors are mounted on the front platform at 60\u00b0 one each other from the central axis, like in Fig. 3. The sensors include the left side, front and right side locations of the robot, namely Left_Sensor (SL), Front_Sensor (SF), Right_Sensor (SR). The detection distance for each sensor is more than 10 cm with 80% sensitivity. After some tests we concluded that IR sensors have a detection range of 10 cm around the mobile robot platform and the detection angle is more about 900 in the front of the robot, like in Fig. 4. The mechanical structure handling this architecture is based on the two DC motors controlling through gears two differential wheels. III. RULE BASED FUZZY NAVIGATION SYSTEM Fuzzy logic unlike classical logic is tolerant to imprecision, uncertainty and partial truth. In the context of mobile robot control, a fuzzy logic based system has the advantage that it allows intuitive nature of sensor-based navigation and can easily transform linguistic information into control signals [15]. The basic structure of a fuzzy logic controller (FLC) consists of three conceptual components: 1) fuzzification of the input\u2013output variables; 2) rule base that contains a set of fuzzy rules; 3) reasoning mechanism that performs the inference proce- dure on the rules and given facts to derive a reasonable output" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000556_j.isatra.2009.05.003-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000556_j.isatra.2009.05.003-Figure1-1.png", + "caption": "Fig. 1. PUMA robot.", + "texts": [ + " Here the uncertainties effect is decomposed as continuous part \u03c4c(q, q\u0307) and discontinuous part \u03c4d(q, q\u0307), u \u2208 Rn vector of joint torques supplied by the actuators; q \u2208 Rn vector of joint positions; q\u0307 \u2208 Rn vector of joint velocities and q\u0308 \u2208 Rn vector of joint accelerations. The control problem can be stated as: for a given bounded desired trajectories qd, q\u0307d and q\u0308d \u2208 Rn design the control input torques \u03c4 such as the robot\u2019s states follow their references, with all involved signals in closed loop remain bounded. lim t\u2192\u221e q(t) = qd(t). A six-degrees-of-freedom PUMA 560 robot is considered for the simulation. Fig. 1 shows the structure of PUMA robot. The kinematical and dynamical parameters of the arm are adopted from the work of Armstrong [21]. The electrical parameters of the motors are assumed by comparing the size and power of the PUMA motors with well-documented commercially-available DC motors [27], and then interpolating the corresponding parameters of interest are taken from our previous work [13]. We assume the robot joints are joined together with revolute joints. Inspired by the idea of basing the fuzzy logic inference procedure on a feedforward network structure, Jang [17] proposed a fuzzy neural network model \u2014 the Adaptive Neural Fuzzy Inference System or semantically equivalently, Adaptive Networkbased Fuzzy Inference System, whose architecture is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000531_9783527627059.ch1-Figure1.32-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000531_9783527627059.ch1-Figure1.32-1.png", + "caption": "Figure 1.32 Magneto-switched dual biosensing of glucose and lactate. (Reprinted with permission from Ref. [164]. 2002 Wiley-VCH Verlag GmbH & Co. KGaA.)", + "texts": [ + " Musameh and Wang [163] also extended the concept of magneto-switching to carbon nanotubes (CNTs) where they took advantage of the magnetic and catalytic properties of CNTs without the need for functionalized magnetic nanoparticles. Willner and coworkers [164] further advanced the utility of magneto-switching by demonstrating the dual quantification of two substrates within a sample. In this work, a gold-coated glass electrode was modified with ferrocene, while magnetic particles were modified with the electron mediator pyrroloquinoline quinone (PQQ) and the cofactor aminoethyl-functionalized-NAD\u00fe (aNAD\u00fe ) as shown in Figure 1.32. Selective quantification of two substrates, glucose and lactate was reversibly accomplished by limiting the potential to a range that was appropriate to activate only one bioelectrocatalytic process. This idea exploits the difference in redox potentials of PQQ and ferrocene. In the case of lactate detection, this is switched on magnetically by bringing the nanoparticles with the cofactor for lactate (NAD\u00fe ) into contact with the electrode. Apotential range of 0.36V to \u00fe 0.15Vwas applied to oxidize PQQ that initiates electron hopping to the electrode when lactate reduces the NAD\u00fe cofactor to NADH in the presence of enzyme lactate dehydrogenase (LDH). The NADH is oxidized by PQQ, which results in the formation of PQQH2 and the regeneration of NAD\u00fe (Figure 1.32a). The electrochemical oxidation of PQQH2 results in the observation of an anodic current.Movement of the magnet to draw magnetic particles away from the electrode, lactate oxidation was blocked and the applied potential was set to allow for only the oxidation and resulting quantification of glucose in the system (Figure 1.32b). Similar to the first example, magneto-switching using ferrocene-modified magnetic nanoparticles are utilized to control the accessibility of mediators to the enzymes by movements of a magnet. A further significant development in the field of magneto-switchable control over magnetic nanostructures on electrodes was the use of alkyl-chain-functionalized hydrophobic magnetic particles in a two-phase system to control and switch the hydrophobic or hydrophilic properties of the electrode surface [165]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001232_cca.2009.5281107-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001232_cca.2009.5281107-Figure1-1.png", + "caption": "Fig. 1. Equivalent formations with time invariant Curvelinear coordinates p, q when following different trajectories.", + "texts": [ + " Less evident, but more important problem for the forma- tion driving of car-like robots is caused by the impossibility to change heading of the robot on the spot. That is why the formations with fixed relative distance in Cartesian coordi- nates cannot be used. To solve this problem, we utilized an approach in which the followers are maintained in relative distance to the leader in curvelinear coordinates with two axes p and q, where p traces \u03a8(t) and q is perpendicular to p as is demonstrated in Figure 1. The positive direction of p is defined from RL(\u03c8L(t)) back to the origin of the movement RL(\u03c8L(t0)) and the positive direction of q is defined in the left half plane in direction of forward movement. The shape of F is then uniquely determined by states \u03c8L(tpi(t)) in travelled distance pi(t) from RL(\u03c8L(t)) along \u2190\u2212 \u03a8L(t) and by offset distance qi(tpi(t)) between \u03c8L(tpi(t)) and Ri(\u03c8i(t)) in perpendicular direction from \u2190\u2212 \u03a8L(t). tpi(t) is the time when the leader was at the travelled distance pi(t) behind the actual position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002736_tmag.2012.2196502-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002736_tmag.2012.2196502-Figure4-1.png", + "caption": "Fig. 4. Analyzed model.", + "texts": [ + " Iron Loss Calculation The iron loss of each finite element, (W/kg) can be calculated directly with the following equation, without the loss data or the approximated function based on measured data where is the core material density and the period of the exciting waveform. The total iron loss is given by (6) where is the thickness of the electrical steel sheet, the number of laminations of the core, and the number of finite elements. Verification of the SCES modeling was performed with a ring-core model. Here we assume that the residual stress distribution in X-direction, is uniform in the ring core and the residual stress in Y-direction, equals zero. Fig. 4 shows the analyzed model and Table I shows conditions used in the analysis, respectively. Fig. 5 shows analyzed results of the maximum magnetic flux density, the maximum magnetic field strength, and the iron loss. As shown in Fig. 5, was large at the inner side and the distribution was changed depending on the assumed residual stress. Because the core material had a weak magnetic anisotropy and the permeability in R.D. was about 1.5 times larger than that in T.D., where the flux was passing in R" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001818_iros.2009.5354819-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001818_iros.2009.5354819-Figure1-1.png", + "caption": "Fig. 1. LineScout motors, cameras and module description.", + "texts": [ + " Although some authors presented control algorithms for the lab prototype of their transmission line robot [9]-[12], the approach described herein was found effective towards the successful introduction of the technology in transmission line maintenance operations. Future applications are briefly discussed. This section presents an overall description of LineScout\u2019s mechanical design, obstacle-crossing strategy, on-board electronics and ground control unit in order to provide a basis for later discussing the control strategy. LineScout\u2019s moving platform was devised as a two-wheel vehicle to optimize energy efficiency as it travels along obstacle-free sections of power lines. Two brushless DC motors, labelled DIS1 and DIS2 in Figure 1, are thus dedicated to LineScout travel. Mounted next to the wheels, a pair of safety rollers, actuated by DC motors ROL1 and ROL2, secures the robot to the line as it stops next to the obstacle to cross. O 978-1-4244-3804-4/09/$25.00 \u00a92009 IEEE 1703 The robot structure that allows the clearing of obstacles is build around three independent frames that slide and pivot relative to one another: the wheels are attached to the wheel frame, the arm frame supports two gripper arms that grasp the conductor to either side of the obstacle while crossing and the centre frame links the first two frames", + " The latter also supports much of the robot\u2019s weight, including the battery and electronics cabinet, plus the two DC motors, ROT and TRA, used to deploy the frames. TRA is equipped with an electromagnetic brake, as are ARM1 and ARM2, which motorize gripper arms GRP1 and GRP2 controlling the opening of the grippers. Lastly, DC motor DPL deploys a mechanism that brings down the wheel frame\u2019s upper portion. The use of these systems is clarified by Figure 2, which presents the crossing of a suspension clamp. Figure 1 also shows the location of the three programmable pan-and-tilt camera (PPTC) units that provide visual feedback to the operator. CAM1 is mounted on the wheel frame, while CAM2 and CAM3 are on the arm frame. The LineARM module, also labelled in Figure 1, is not required to cross obstacles. However, when LineScout is equipped with this module, it allows a fourth PPTC and various tool modules to be used. The LineScout communication scheme revolves around two major systems: the on-board electronics (OBE) installed inside the robot, and the ground control unit (GCU) used by the operator. OBE subsystems are now described using Figure 3. The bidirectional data antenna (1) is linked to the RF card (2), which communicates with the GCU. The converted radio signal is carried over an RS-232 link to the main controller (3), which interfaces the RS-485 bus (4)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002148_j.triboint.2012.03.013-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002148_j.triboint.2012.03.013-Figure3-1.png", + "caption": "Fig. 3. Schematic of hypoid final drive unit showing main components.", + "texts": [ + " The proposed churning loss formula for the hypoid axle from [18] was used here along with the consideration for a pinion from [17]. The gear churning loss was about 800 W at 760 rev/min (crownwheel speed) for the gear oil and depended on (speed)1.31. Tch \u00bc 2:186 hd Rp 0:147 Vo Rp 0:198 Re 0:25 Fr 0:53 \u00f01\u00de The axle contains eight bearings, two taper roller bearings supporting the input pinion, two more supporting the crownwheel, carrier and differential assembly and two parallel roller bearings on the outboard end of each of the two half-shafts only one of which is in contact with the axle lubricant (Fig. 3). Since these outer bearings are physically remote from the main drive unit they are considered thermally distinct and their contribution was neglected. The seal losses were neglected. Bearing losses were computed using the method described in the SKF catalogue [19] with the empirical constants derived from those published by the company for similar bearing types and sizes to those used in the vehicle. Both assembly pre-load and service gear loading were included. Table 4 gives bearing details. The total power loss is, by definition, the sum of the individual componentsP P\u00bc (Gear friction power loss\u00fegear churning power loss\u00febearing power losses) An overall, time dependent, efficiency, Z, is then defined as Z\u00bc 1 SP=P\u00fe1 \u00f02\u00de where P is the instantaneous output power" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001034_j.triboint.2009.08.005-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001034_j.triboint.2009.08.005-Figure1-1.png", + "caption": "Fig. 1. The contact schematic diagram of a rigid flat with an ellipsoid.", + "texts": [ + " The inception of the elastoplastic deformation regime of an ellipsoid is determined using the theoretical model developed for the yielding of an elliptical contact area. However, the inception of the fully plastic deformation regime is solely determined by the finite element analysis such that the contact pressure, P, and the equivalent von-Mises stress values, (Seqv), formed at the contact area are uniformly distributed. ke is defined as the ellipticity of an ellipsoid, so the contact parameters shown in the elastoplastic deformation regime are evaluated by varying the ke value. In the present study, Fig. 1 shows that a deformable ellipsoid tip contacts with a rigid flat. The lengths of the semi-major axis of an ellipsoid and the semi-minor axis are assumed to be cR\u00f01rco1\u00de and R, respectively. From the geometrical analysis, the radii of curvature at the tip of an ellipsoid, R1x\u00f0 \u00bc c2R\u00de and R1z\u00f0 \u00bc R\u00de, are obtained. The ellipticity of an ellipsoid is defined as ke, and ke \u00bc \u00f0R1z=R1x\u00de 1=2 \u00bc 1=c\u00bc R=cR. For c=1, ke=1 corresponds to the spherical contact; for c-1, ke=0 corresponds to the cylindrical contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002760_s00542-010-1188-4-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002760_s00542-010-1188-4-Figure5-1.png", + "caption": "Fig. 5 Coordinate system for the tilting motion of the rotating part", + "texts": [ + " This model consists of two grooved journal bearings, four plain journal bearings, one grooved thrust bearing, and one plain thrust bearing. The fluid film was discretized by 7,000 four-node isoparametric bilinear elements. The accuracy of the developed program was verified by comparing the calculated flying height of the coupled journal and thrust bearings in equilibrium (where the axial load generated by the FDBs was equal to the weight of a rotor) with the measured flying height at various rotating speeds (Jang et al. 2006). Figure 5 shows the coordinate system for the tilting motion of the rotating part. The tilting ratio is defined as follows: g \u00bc hl 1 2 L \u00f022\u00de where L and hl are the length of the journal bearing and the distance between the bearing span center, R0, and the tilting center, R1, respectively. If the tilting center was ?Z from the bearing center, the tilting ratio had a positive sign. Otherwise, it had a negative sign. Figures 6 and 7 show the variation in the stiffness and damping coefficients of the FDBs due to a change in the eccentricity ratio" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001074_j.triboint.2010.10.002-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001074_j.triboint.2010.10.002-Figure4-1.png", + "caption": "Fig. 4. Reinforcement of the polymer compression platens for the ring tests. (a) Support block and shrink ring. (b) Upper and lower PC compression platens mounted on the shrink-fit set-up. (c) Schematic representation of the shrink-fit set-up of the lower compression platen.", + "texts": [ + " The desired texture and roughness of the surfaces of the compression platens were produced and regenerated after completion of each test by means of a grinding and polishing procedure that is schematically shown in Fig. 3. The procedure combines the utilization of sand paper or polishing cloths with the rotation of the compression platen in order to ensure predominant radial patterns and average values of surface roughness (Ra) in the range 0.04\u2013 0.65 mm (Fig. 3b). These values were measured by a roughness tester along the radial direction of the platens. The ring test specimens were centred between the platens and compression was performed to a pre-determined height reduction (Fig. 4). Inner diameters and heights of the specimens were measured and recorded before and after the tests. The final values of these quantities were determined as the average of three measurements of internal diameter and height utilising a digital caliper. In case of the platens made from PVC and PC, the maximum applied pressure during the ring compression of lead was found to be close to the yield strength of the polymers (Table 1). This influences the mechanical response of the platens and material flow during the ring compression test", + " As a result of this, some experiments were performed with a modified tool design that inserts the compression platens in a shrink ring to provide reinforcement. The resulting tool set-up prestresses the platens and ensures a firm external support that prevents its radial deformation under loading. The interference between the ring and platen is controlled by means of a conical press fit (angle of inclination equal to 51) that incorporates tightening screws to limit and control the overall level of pressfitting (Fig. 4). However, in order to understand the influence of the aforementioned reinforcement additional ring tests were performed in PVC compression platens without press-fitting. Table 2 presents the experimental workplan for the ring compression tests. The experiments were designed in order to isolate the influence of process parameters that were considered more relevant for the overall investigation on frictional behaviour: (i) surface roughness of the tools and (ii) differences in material strength of the tribo-pairs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002702_cdc.2011.6161241-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002702_cdc.2011.6161241-Figure2-1.png", + "caption": "Fig. 2. Bicycle model.", + "texts": [ + " However, with suitable modifications, the basic bicycle model can be modified to incorporate the effects of longitudinal load transfer ([11], [12]) and the friction circle constraint ([11], [12], [13]). In this work, longitudinal load transfer is not included, but the friction circle constraint is modeled, due to the nonlinear tire response encountered for high tire slip angles arising from aggressive yawing or drifting. The bicycle model, with relevant nomenclature and conventions, is shown in Fig. 2. The vehicle\u2019s state vector is given by x = [u, v, r, \u03c8, x, y]T , where u, v are the components of the vehicle\u2019s total velocity V along the body-fixed xb and yb directions, i.e. V = \u221a u2 + v2, r is the yaw rate, \u03c8 is the heading, and x, y are the coordinates of the vehicle\u2019s CG measured from a fixed origin. The control vector is chosen as u = [\u03b4, Fxf , Fxr,Md] T , featuring, respectively, the steering angle, the front and rear tire longitudinal forces and the direct yawing moment generated using TV", + " Md is the direct yawing moment created as a result of TV. Fi\u2217 (i=x, y; \u2217=f, r) denote the longitudinal and lateral forces developed at the front and rear tires, referred to a tire-fixed reference frame. These forces depend on the normal loads on the front and rear axles, Fzf and Fzr . For this analysis, the latter are treated as equal to their steady-state (static) values as follows, Fzf = mg\u2113r \u2113f + \u2113r , Fzr = mg\u2113f \u2113f + \u2113r , (2) where \u2113f and \u2113r are the distances of the front and rear axles from the CG of the vehicle respectively, as shown in Fig. 2. Since the aggressive maneuver analyzed in this paper involves large slip angles, the linear force-slip relationships, such as those found in [14], [15] are not applicable. Third order polynomial [16] and rational function approximations [17] can capture the nonlinear tire dynamics and saturation effects, but in this work, the well-known Pacejka \u201cMagic Formula\u201d (MF) [18] is used instead. The MF models the tire forces as transcendental functions of the tire longitudinal and lateral slips. Using this representation, the lateral tire forces are expressed as Fy\u2217 = \u2212Fmax y\u2217 sin(C arctan(Bsy\u2217)), \u2217 = f, r (3) where the constants B and C depend on the tire and road surface characteristics, and syf and syr are the lateral slip ratios of the front and rear tires, expressed as syf = v cos \u03b4 \u2212 u sin \u03b4 + r\u2113f cos \u03b4 u cos \u03b4 + v sin \u03b4 + r\u2113f sin \u03b4 , (4a) syr = v \u2212 \u2113rr u ", + " It should be noted that during the course of the rotation, the front tire is always driven according to a fixed ratio to the driven rear tire (left or right), and never braked, i.e., Fxf \u2265 0. Assuming a clockwise rotation as viewed from above, the pro-steering phase involves a steering deflection to the right (\u03b4 > 0). Since initially there is no lateral velocity or yaw rate, this gives Fyf \u2265 0. If at this time the front wheel was braked, (i.e., if Fxf < 0), it is clear from the vehicle geometry (Fig. 2) that the moment due to Fxf would have opposed that due to Fyf . In other words, braking would have partly negated the moment generated through steering input, and thus adversely affected the yaw acceleration. For the yaw deceleration/arrest phase, counter-steering implies \u03b4 < 0 which, depending on the vehicle\u2019s lateral velocity and yaw rate, may yield Fyf < 0. In this case also, the moment generated by braking the front tire opposes that generated by the lateral force of the front tire, Fyf . For this reason, the condition Fxf \u2265 0 is maintained for the duration of the maneuver" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure2-1.png", + "caption": "Figure 2: Variator Force Balance", + "texts": [ + " When the rotational velocities of the input and/or output discs change, the rollers automatically alter their inclination in order to adjust to the new operating conditions (4). Power transmission is achieved by traction, i.e. by shearing an extremely thin, elasto-hydrodynamic fluid film (traction fluid [1]) and not through metal-to-metal friction. Hence the name 'traction drive', which is defined in [2] as: \u201ca power transmission device which utilizes hardened, metallic, rolling bodies for transmission of power through an elastohydrodynamic fluid film\u201d. The force balance in the discs and rollers is shown in figure 2. The application of a castor angle to the roller carriages (as described in Figure 2) enables the rollers to \u2018steer\u2019 to a new angle of inclination and hence Variator ratio. The Torotrak Variator is torque controlled in that the required system torque is set by applying pressure to the pistons connected to the rollers and the Variator follows the ratio automatically [3]. Figure 3 explains this approach using a simplified single roller model. Applying a reaction force F to the roller causes a reaction torque (Ta and Tb) at the Variator discs and consequently an acceleration of the two inertias (engine side inertia A and vehicle side inertia B)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003525_tmag.2010.2089501-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003525_tmag.2010.2089501-Figure3-1.png", + "caption": "Fig. 3. Combination of main 2-D FEM and post 1-D FEM.", + "texts": [ + " In this section, the line-line voltage is assumed to be sinusoidal wave in order to focus on the effect of the slot harmonics. Fig. 2 shows the 3-D finite-element mesh for the direct consideration of the eddy currents in the core. The tetrahedral- edge finite elements are used for discretization. The analysis region is reduced to half the thickness of the steel sheet by imposing boundary conditions [3]. In this case, the eddy-current loss in the core is considered in (1), whereas the hysteresis loss is still neglected. Therefore, the torque obtained by this analysis is compensated by (9) without . Fig. 3 shows the meshes for the proposed method. The 2-D approximation is applied for the main FEM. In the post 1-D FEM, half the thickness of the electrical steel sheet is subdivided into six elements. Shorter elements are applied for the surface side due to the skin depth. Fig. 4 shows the calculated iron loss including the harmonic secondary cage loss at no-load conditions. The results obtained by the 2-D FEM without the post 1-D FEM are also shown. It is observed that these results are nearly identical at the low speed, whereas the result obtained by the 2-D FEM overestimates the loss at the high speed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001554_te.2009.2027333-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001554_te.2009.2027333-Figure2-1.png", + "caption": "Fig. 2. A slide from the mechatronics lecture explaining the relationship between torque , angular speed , and the radius of two gears.", + "texts": [ + " The lecture introduced the use of a gearbox and discussed how each wheel of a two-wheel robot can be driven independently using two dc motors and a gearbox. The motors used in the gearbox are high-speed/ low-torque motors, and the need for a gearbox to increase the output torque in order to overcome the friction of the surface the robot ran on was introduced through the idea of levers, wrenches, and forces. It was expected that students had already seen these physical concepts in their high school physics courses. This discussion prompted an introduction to gear ratios and how the gear ratio relates to angular velocity, force, and torque. Fig. 2 shows one of the slides from the lecture explaining the relationship between torque, angular speed, and gear radius. The lecture also introduced the idea of feedback control using sensors. The fourth author described how a person goes about picking up an object by using their eyes as optical sensors and their hands as tactile sensors to feel when the object is firmly grasped. The lecture correlated this process with the use of sensors in mechatronics systems, using the example of a simple switch that can be either pressed or not to indicate the state of the system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000465_iet-smt:20060018-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000465_iet-smt:20060018-Figure2-1.png", + "caption": "Fig. 2 Eigenmode 2", + "texts": [ + " The stator of the IM has a diameter of 140 mm, and its yoke has a thickness of approximately 10 mm. The flux is the rated flux, and there is no load. Table 1 shows the results of the eigenmode analysis for the most important eigenmodes of the two-pole machine. As the current has a frequency of 50 Hz, the main vibration frequency is 100 Hz. Therefore, only the data for this frequency are given. It is obvious that the second order modes, that is modes 2 and 20, are the most important modes for the two-pole machine. In Fig. 2, vibration mode 2 is shown. The contour of the undeformed machine hull is represented by the black line. Mode 20 has the same shape as mode 2, but is shifted over 458. The generalised vibrations of mode 2 and 20 at 100 Hz have a phase shift of 908. The combination of the vibrations of modes 2 and 20 produces rotating waves. Inspection of the results for mode 2 in Table 1 leads to the conclusion that, for this mode, the contributions of the magnetic forces and the magnetostriction to the vibrations are in phase and thus add up" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003312_j.jsv.2013.05.022-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003312_j.jsv.2013.05.022-Figure5-1.png", + "caption": "Fig. 5. Open-circuit magnetic field distributions, where (a) 1-magnet/12-slot, (b) 8-magnet/1-slot, and (c) 8-magnet/12-slot.", + "texts": [ + " The radial and tangential magnetic flux densities are presented to further clarify the effect of magnet/slot combination on magnetic forces, as shown in Fig. 4. The results show that the waveforms in the 1-magnet/15-slot and 8-magnet/1-slot motors exhibit weak symmetry, which repeat themselves after a complete rotation as the magnet number is prime to slot number. By contrast, the flux density in the 8-magnet/15-slot motor has more symmetry because its periodicity is more significant. Consequently, the magnetic forces in the latter exhibit more self-equilibration. Fig. 5 illustrates the field distributions of the even-slot motor. Similar to the odd-slot one, the distributions in the 1-magnet/12-slot and 8-magnet/1-slot motors exhibit weak symmetry (Fig. 5(a) and (b)), whereas more symmetry can be found in Fig. 5(c) in the 8-magnet/12-slot motor. Thus one can predict that the forces in Fig. 5(c) are balanced well and create less UMF than others. Fig. 6 shows the radial and tangential magnetic flux densities within the air gaps, where the waveforms of the 1-magnet/ 12-slot and 8-magnet/1-slot motors are 2\u03c0-period and exhibit weak symmetry because the magnet number is prime to slot number. The waveform of the 8-magnet/12-slot motor has more symmetry due to the significant periodicity, and basically there are always four identical sections in this motor no matter what the relative positions between the rotor and stator are, and thus the radial forces can be naturally canceled", + " In contrast, the 24th order is definitely excited because it is the equivalent 3rd order of the magnet-frequency or the 2nd order of the slot-frequency forces, where the factors are km\u00bcmod(lmNm/Ns)\u00bcmod(3 8/12)\u00bc0 and ks\u00bcmod(lsNs/Nm)\u00bcmod(2 12/8)\u00bc0, respectively. 3.3.3.1. SHAFT-FREQUENCY EXCITATION ANALYSIS. When considering the shaft frequency force, quite a different situation arises, thought the occurrence or suppression behaviors of the magnetic forces can also be addressed in a similarly way. The 1-magnet/15-slot motor in Fig. 3 and 1-magnet/12-slot motor in Fig. 5 are equivalent to rotors with moving eccentricity, where as the 8-magnet/1-slot motors in Figs. 3 and 5 can model the stators with eccentricity, where the eccentricity is movable with the rotating stator for the odd-slot motor but unmoved for the even-slot motor. 3.3.3.2. HARMONIC FORCE IN THE 8-MAGNET/15-SLOT MOTOR. Physically, the 1st, 14th, 29th, and 31st orders in Fig. 7, (c) and (d) are all related to the shaft-frequency for the 1-magnet/15-slot motor. These are remarkable because they are in essence the 0th, 15th, and 30th harmonics, where the 0th harmonic corresponds to the direct component in the force on the magnet causing UMF and CT, and the rest two have the factors km0 \u00bcmod(lm/Ns)\u00bc{mod(15/15), mod(30/15)}\u00bc{0, 0}, also implying excited UMF and CT" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003712_01691864.2015.1023219-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003712_01691864.2015.1023219-Figure1-1.png", + "caption": "Figure 1. Symmetric spatial hybrid parallel manipulator.", + "texts": [ + " Practically, the degree-of-freedom value F = 12 of the hybrid mechanism is equal to the degrees of freedom associated with all the moving links \u03bd = 72 minus the total number of independent constraint relations l = 60 imposed by the joints. The two moving platforms are initially considered at a central configuration, where are not rotated with respect to the fixed frame and their centres are located at given relative elevations OG = GH = h above the origin O of fixed base. For the purpose of analysis, we assign a fixed Cartesian coordinate system Ox0y0z0 at the point O of the fixed frame and also two appropriate mobile frames GxGyGzG and HxHyHzH on the moving platforms at their mass centres G and H, respectively (Figure 1). The active leg A of lower module, for example, consists of a little cross of a universal joint of mass m1linked at the frame A1xA1y A 1 z A 1 which has the angular velocity xA 10 \u00bc _uA 10 and the angular acceleration eA10 \u00bc _xA 10 connected at a moving cylinder A2xA2y A 2 z A 2 of length l2, mass m2 and tensor of inertia J\u03022, having a relative rotation around A2zA2 axis with the angle uA 21, so that xA 21 \u00bc _uA 21, e A 21 \u00bc _xA 21. An actuated prismatic joint is as well as a piston of length l3, mass m3 and tensor of inertia J\u03023, linked to the A3xA3y A 3 z A 3 frame, having a relative velocity vA32 \u00bc _k A 32 and acceleration cA32 \u00bc _vA32" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002229_tpas.1971.292930-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002229_tpas.1971.292930-Figure5-1.png", + "caption": "Fig. 5. Shell of finite thickness and infinite length. Current filament outside shell.", + "texts": [ + "9 the current density due to combination of proximity effect and skin effect at three different radii, is shown in Fig. 2. The ordinate of these curves is the magnitude of current density in amperes/cm2 while the abscissa is the angle from 0 to 7r. Fig. 3 shows the corresponding curves of the phase difference y. Three curves 1, 2 and 3 of current density and angle y are shown corresponding to radii r= 0.100m - 0.106m and 0.1 12m. CYLINDRICAL SHELL, CURRENT OUTSIDE Development of Equations The current filament is placed at a distance b from the axis of the shell, Fig. 5. The eddy current density distribution in the material of the shell i(r,o) is given by Eq. 1 and the magnetic vector potential within the material Ao(r,k) is given by Eq. 2. Inside the shell, the magnetic vector potential Aa, i (p,G) will be to the eddy current distribution in the shell. A0 i(p,O) is derived in a similar way as in the case when the current was inside the shell, and is given by Eq. 6. Outside the shell, the magnetic vector potential Aa,e(P,0) due to the eddy current distribution is derived again in a similar way as in the case when the current was inside the shell and is given by Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002159_pime_proc_1969_184_061_02-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002159_pime_proc_1969_184_061_02-Figure6-1.png", + "caption": "Fig. 6. Elemental roller", + "texts": [ + " Pressure at which value of pressure-viscosity expo- Absolute viscosity, reyns, lb s/in2. Equivalent viscosity in the contact ellipse, reyns, Equivalent viscosity at y = b, reyns, lb s/in2. Absolute viscosity at atmospheric pressure and tem- nent changes, lb/in2. lb s/in2. perature, reyns, lb s/in2. Fig. 5 is a representative contact ellipse for a ball in a nonconforming groove. The friction force M, about the z axis due to the section of the ellipse for y > b and y < - b is obtained by integrating the product of the friction of the elemental roller (Fig. 6) and the moment arm (10) thus, - 0 M I = 21 y d F . . . (42) t h Proc lnstn Mech Engrs 1969-70 Vol 184 P t 1 No 44 at UNIV CALIFORNIA SAN DIEGO on June 21, 2016pme.sagepub.comDownloaded from 846 COMMUNICATION - N Per cent con fo rm i t y The above can be numerically integrated. There remains the area within the inscribed circle where a true elastohydrodynamic film theory is impossible to maintain. Therefore, the film thickness h is assumed to be the same as that at y = b. If the inscribed circle is then divided into elemental rings of radius r and width dr, from (10)\u2019 the torque M, for the inscribed circle becomes M, = /ob r3pb dr " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000870_065102-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000870_065102-Figure15-1.png", + "caption": "Figure 15. Experimental setup 2.", + "texts": [ + " Figure 14(b) displays the envelope spectrum of the bandpass filtered signal. The effectiveness of the proposed fault detection approach can be evidenced by comparing figures 13 and 14(b). One would appreciate the proposed method by noting that such precisely designed bandpass filters are rarely available in practice, particularly for online monitoring applications. 3.4.2. Detecting inner race fault of an ac motor bearing. In this test, an ac motor with a faulty bearing is installed on the simulator as shown in figure 15. The ac motor bearing is of type NSK-6203 (inside, outside, pitch and ball diameters are, respectively, 0.6693\u2032\u2032, 1.5748\u2032\u2032, 1.142\u2032\u2032 and 0.266\u2032\u2032) with eight rolling elements (balls). A single balanced load rotor (2\u2032\u2032 thick, 4\u2032\u2032 in diameter and 11.1 lbs) is mounted on the same 5/8\u2032\u2032 steel shaft used in the previous test. Both bearings that support the shaft are normal. To introduce additional vibration interferences, two unbalanced rotors (0.5\u2032\u2032 thick, 6\u2032\u2032 in diameter and 1.42 lb each, with two 0", + "011 lb unbalancing weights mounted to the outer perimeter of each disk) are also installed on the 5/8\u2032\u2032 steel shaft. Similar to the previous experiment, the gearbox is driven by the shaft via a belt connection. The shaft speed is set at 1428 RPM (23.8 Hz). The ac motor bearing has a pre-seeded fault (created by the manufacturer with unknown dimensions) on the inner race with a characteristic frequency of 117 Hz (=4.932 fr , specified in the simulator user\u2019s manual). The accelerometer is installed on the simulator base and away from the ac motor as shown in figure 15. Figure 16 displays a portion of the measured data and their frequency domain representation. A close look at the spectrum of the measured signal (figure 16(b)) reveals a region (1550\u20131900 Hz) of high SNR associated with the excited resonance frequency. The envelope spectrum of the measured signal is shown in figure 17. As seen from figures 16(b) and 17, the frequency components associated with the interfering vibrations at gear meshing, 167.3 Hz, and associated harmonics as well as the harmonics of the shaft rotational frequency (23" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000466_iros.2008.4650802-Figure17-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000466_iros.2008.4650802-Figure17-1.png", + "caption": "Fig. 17. Overall system for half hitch production.", + "texts": [ + " 16 and Fig 18 show sequences of continuous photographs of the knotting tasks (overhand knot and half hitch). The knotting strategy used was the one proposed in a previous paper [1]. Overhand knot The experimental system is shown in Fig. 15. Fig. 16(a) \u223c (e) show loop production. In Fig. 16(f), the rope sections are pressed by the free finger to strengthen the contact state between the two sections. Fig. 16(g) \u223c (i) show rope permutation. Fig. 16(j) \u223c (l) show rope pulling. Half hitch The experimental system is shown in Fig. 17. In the initial state, the rope is wrapped around the object, as shown in Fig. 17. Fig. 18(a) \u223c (c) show loop production. In Fig. 18(d), the rope sections are pressed by the free finger to strengthen the contact state between the two sections. Fig. 18(e) \u223c (g) show rope permutation. Fig. 18(h) \u223c (i) show rope pulling. Finally, Fig. 18(j) \u223c (l) show additional rope pulling by a human hand to tighten up the knot. These video sequences can be viewed on our web site [7]. First, to identify the necessary skills for knotting, we analyzed a knotting action performed by a human subject" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000445_13552540810841562-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000445_13552540810841562-Figure2-1.png", + "caption": "Figure 2 Schematic diagram of HPDM", + "texts": [ + " The flowchart of HPDM direct rapid manufacturing technology shown in Figure 1 is based on the layered deposition of molten powder material using plasma arc, which is one of most economic ways of depositing molten metals. The STL file of the 3D CAD model of product is imported into the HPDM software system, then the processing path and NC instructions are automatically generated by the software. Based on these, the combination of plasma deposition and NC milling is realized by the hybrid machine. The schematic diagram of the HPDM machine is shown in Figure 2. It has two working procedures: deposition by plasma torch shown on the left and surface finishing by NC milling head on the right, respectively. In the HPDM process, when a layer is deposited, its top surface is machined used planar milling to obtain a smooth surface with a certain thickness for further deposition, and the internal and external surface profile is machined used contour milling with T-slot cutters to remove the remaining stair steps on the surface and to attain fine surface state of the near-net shape metal part" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002940_05698196508972108-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002940_05698196508972108-Figure4-1.png", + "caption": "FIG. 4. Varia tion of the mixing length upon the normal to the lubricating film.", + "texts": [ + ",f h'~ 1 f/ /t j v/ tt/j ~ :0vv~ '1 IV-I 1/ k:~ ~v Jr;1/ ./ / / t:Y v ..... V Ii 1/ 1/ I l/ V i/ I /./ Vy I J I 7:0 I -'W, I I I -20 -30 - 30 20 30 ~ ~ c .~ - 10 c; Cl.> E zs > E d l.... & 0 F r\u00ab. 2. Fricti on st resses To, T h 0 11 t he two lubrica ted sur faces ve rsus B, for variou s modified R eyn old s numbers R. Indeed , in T abl e 2 the va lues Tc calcula te d numeri cally (Table 1) and by means of formula [27] are presented. The sa me results are graphicall y represented in Fig. 4. It is interesting to point out that relation [27] agrees for R = \u00b0 (laminar regime) . T he numerica l values obtained for Tc are given in F ig. 3. It may be seen that rela ti on [26] gives a good accuracy for 100 < R < 1,000 if K; = 0.0525 and n c = 0.75, that is -Tc = 1 + 0,052.5 R o.-i5. [27] with the known results of fluid mechanics, that is, for rela ti vely grea t Reynolds numbers t he fri ction stress is T A B L E 2 Th e St ress Te fo r Various Reynolds N umb ers R 100 200 500 1,000 ITel - 1 numericall y ('f able 1) 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure10-1.png", + "caption": "Fig. 10. Comparison of optimized rotor profile with \u2018\u2018FS\u201d rotor profile.", + "texts": [ + " An experiment was conducted to verify that the proposed rack-generated rotor profiles have the excellent performance. The optimized (rack-generated) and well-known \u2018\u2018FS\u201d (Fu\u2013Sheng) [14] rotors are compared within the same oil-flooded air compressor casing and under the same working conditions. The ratio of screw length to outer diameter of the male rotor was fixed as 1.62, the wrap angle and helix angle were fixed as 299.622 and 46.0 respectively. The optimized and \u2018\u2018FS\u201d rotor profiles used in the experiments can be seen in Fig. 10. In the experiments, three pairs of finished rotors with the optimized profile were tested from 1800 to 4800 rpm with interval 600 rpm. The pressure and temperature were 1.017 kg/cm2 and 24.4 C for the air inlet, and the pressure of air outlet was 7.01 kg/cm2 in average. The temperature of the oil inlet was 54.5 C and the oil flow rate was increased from 18.09 to 49.86 lit/min with the motor speed. Moreover, the more smooth and exact discharge channels, smaller interlobe-, end- and tip-clearances were carefully manufactured in accordance with the optimized rotor profile and the build-in volume ratio" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002789_cjme.2013.03.532-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002789_cjme.2013.03.532-Figure3-1.png", + "caption": "Fig. 3. Geometric relationships considering overturning angular disturbance", + "texts": [ + " The cutter head of the TBM often endures huge overturning moment and it will bring the overturning angular displacement disturbance which cannot be ignored. Being affected by gear radial displacement disturbance and overturning moment, it will lead to four changes as follows. Supposing the change value of center distance is L\u2206 , actual center distance is 12F and original center distance is L12, then the geometrical relationship between the actual center distance and original center distance is shown in Fig. 3(a). In Fig. 3(a), 12F is the meshing force between the pinion and ring gear. The actual center distance and change value of center distance can respectively be expressed as 2 2 1 2 12 12 2 1 2 1( ) ( ) ,O O L L x x y y (4) 12 12 2 1,L L L x x\u2206 (5) where x2 and x1 represent the coordinate values in x direction of the two gears, respectively. The relationship between the center distance and meshing angle of internal meshing gears is br bp 0 12 cos r r L \u03b1 . (6) The angular increment \u03b1\u2206 of original meshing angle \u03b10 owing to the change of center distance takes first order approximation of the Taylor\u2019s expansion, which can be expressed as br bp 2 1 2 1 2 12 012 0 ( )( ) tancos r r x x x x LL \u03b1 \u03b1\u03b1 \u2206 . (7) Taking the vector of the overturning angular displacement to project along the meshing line, the relationship between the radial linear displacement and overturning angle is shown in Fig. 3(b). Then \u03bac in plane yO2z centered at O2 and \u03bac can be expressed as c , 0 , 0sin cosx i y i\u03ba \u03ba \u03b1 \u03ba \u03b1 . (8) Similarly, the original meshing angle \u03b10 in plane 2yO z , is centered at 2O . The actual meshing angle\u03b1 affected by the overturning angular displacement disturbance \u03bac and \u03b10 meet the space law of cosines. According to the space law of cosines, the relationship among\u03b1 , \u03bac and \u03b10 is c 0cos cos cos\u03b1 \u03ba \u03b1 . (9) Therefore, the meshing angle change \u03b1\u2206 under overturning angular displacement disturbance is c 0 0arccos(cos cos ) " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003953_15325008.2014.880968-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003953_15325008.2014.880968-Figure1-1.png", + "caption": "FIGURE 1. Cross-sectional geometry of novel BSRM.", + "texts": [ + " The rotor eccentricity greatly affects the suspending force, because the non-uniform air gaps between the rotor and stator tooth poles cause imbalanced radial net force. First, the new BSRM\u2019s air-gap permeances varying along with the rotor eccentricity are analyzed. The radial force model describing the relationship between the radial force, winding currents, rotor position angle, and rotor eccentricity displacement is then derived. Finally, the Finite Element Method (FEM) results verify the proposed radial force model. The novel BSRM is shown in Figure 1 [15], where the torque main windings Na, Nb, and Nc for phases A, B, and C, respectively, are the same as that of the conventional BSRM, but there are only two sets of suspension windings Nx and Ny, rather than six sets of suspension windings Nax, Nbx, Ncx, Nay, Nby, and Ncy in the conventional BSRM shown in Figure 2 [2]. The adjacent three stator teeth of phases A, B, and C share the X -axis suspension winding Nx, consisting of two coils connected in series on the face-to-face (or diametrically opposed) poles", + " The BSRM\u2019s three-phase torque windings are energized in turn, and there is only one phase torque main winding being energized at one moment. When the B phase torque main winding Nb and suspension windings Nx and Ny are energized with currents imb, ix, and iy, respectively, the magnetic equivalent circuit is shown in Figure 3, where Pb1\u2013Pb8 are the permeances D ow nl oa de d by [ U ni ve rs ity o f St el le nb os ch ] at 0 3: 26 0 5 N ov em be r 20 14 of the air gap between the stator and rotor tooth poles, \u03c6b1\u2013\u03c6b8 are the magnetic fluxes, and subscripts 1, 2, . . . , 8 denote the reference number of the stator poles shown in Figure 1. The mutual inductances between the three torque main windings can be ignored [17]; Pb9, Pb10, Pb11, and Pb12 are assumed as zero. The rotor rotates clockwise from the initial position \u03b8 = 0\u25e6 defined in Figure 4. The rotor position range is from 0\u25e6 to 15\u25e6 for the B phase torque main winding being energized. From Figure 3, one obtains\u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03c6b1 Pb1 + Nmimb + Nxix = \u03c6b2 Pb2 \u2212 Nmimb + Nyiy = \u03c6b3 Pb3 + Nmimb \u2212 Nxix = \u03c6b4 Pb4 \u2212 Nmimb \u2212 Nyiy, \u03c6b5 Pb5 + Nxix = \u03c6b6 Pb6 + Nyiy = \u03c6b7 Pb7 \u2212 Nxix = \u03c6b8 Pb8 \u2212 Nyiy = \u03c6b1 Pb1 + Nmimb + Nxix " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000633_1.3142868-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000633_1.3142868-Figure4-1.png", + "caption": "Fig. 4 Basic mappings: \u201ea\u2026 impacting chatter only and \u201eb\u2026 with stick switching. The straight lines with arrow represent impacts on the boundaries.", + "texts": [ + " From the stick switching lanes, the mappings are defined as P1: 21 i \u2192 12 i , P2: 12 i \u2192 21 i , P3: 12 i \u2192 23 i P4: 23 i \u2192 32 i , P5: 23 i \u2192 32 i , P6: 32 i \u2192 21 i 19 ith mixed switching planes, four mappings are defined by P2: 12 i \u2192 R 2 i , P2:R 2 i \u2192 21 i P3: 12 i \u2192 L 2 i , P3:R 2 i \u2192 23 i P5: 23 i \u2192 L 2 i , P5:L 2 i \u2192 32 i P6: 32 i \u2192 R 2 i , P6:L 2 i \u2192 21 i 20 mong four basic mappings, the two mappings P2 and P5 are ocal and the other two mappings P3 and P6 are global. The local apping will map the motion from a switching plane onto itself. owever, the global mapping will map the motion from a switchng plane to another one. Such mappings are sketched in Fig. 4 a . he corresponding switching planes and mappings are labeled. On he impacting chatter boundaries, impacts are expressed by thin traight lines with arrows. The mappings relative to the stick witching planes only are sketched in Fig. 4 b . Only two stick appings P1 and P4 are new, and the other four mappings are he same as in Fig. 4 a . The mappings based on the sticking and mpacting switching planes are presented in Figs. 5 a and 5 b . Set a vector as 41013-4 / Vol. 131, AUGUST 2009 om: http://vibrationacoustics.asmedigitalcollection.asme.org/pdfaccess.as yk tk,xk i , x\u0307k i , x\u0307k i\u0304 T 21 For the impacting maps P =1,2 , . . . ,6 , yk+1= P yk can be expressed by P : tk,xk i , x\u0307k i , x\u0307k i\u0304 \u2192 tk+1,xk+1 i , x\u0307k+1 i , x\u0307k+1 i\u0304 22 From Appendix or in Ref. 37 , the absolute displacement and velocity for two gear oscillators can be obtained with initial con- ditions tk ,xk i , x\u0307k i and tk ,xk i\u0304 , x\u0307k i\u0304 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001466_s0019-9958(74)90833-x-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001466_s0019-9958(74)90833-x-Figure4-1.png", + "caption": "FIG. 4.", + "texts": [ + " At the beginning of the first standard stage there will be 31 cells, three of which are markers, dividing the filament into two equal subsegments, and the length of each subsegment (including the adjacent marker cells) will be 15. Let us assume that we are at the beginning of a standard stage. That is, the two end cells are markers, the number of consecutive markers is not more than two anywhere in the filament, and the length of each subsegment is everywhere the same. Let l denote this distance (see Fig. 4). Then the number of consecutive nonmarker cells is 1 - - 1. Because one may find markers both singly and in pairs, it is easier to think of the segments between the markers as the primary concept, and to talk about the end-markers of a segment. A marker may be the end-marker Standard stage in the synchronization of a growing filament. of one or two segments. In what follows, we shall talk about the waves initiated by the left end-marker and the right end-marker of a certain segment. I f a cell is an end-marker for two segments, it simultaneously generates waves for both segments. Consider the right end segment of the filament after time wl, where w is the least common multiple of q and t, such that its leftmost cell is the original left end-marker of the rightmost subsegment. The length of this segment is l[(pw/q) + 1]. Since this is a multiple of l, it is possible to divide it into equal subsegments of length [//2]. The number of these subsegments will be m = 2[(pw/q) + 1]. Note that m does not depend on I. Consider the points (marked 0, 1, 2,..., m in Fig. 4) which divide the segment under consideration into m equal parts. Since l is an integer and the distance between two consecutive points of this type is 1/2, these points fall either exactly in the center of a cell or exactly on the boundary of two cells. We make our model such that, if one of these points is in the middle of a cell, then that cell becomes a marker, and if the point is on the boundary of two cells, both cells become markers. To show that this results in subsegments where the distance between the end-markers is [//2], we distinguish between the cases when I is even and when 1 is odd", + " Note that, since w >~ q > p, and p a n d q are integers, we have lzJ ] ~< 1 and ]z~,l ~ 1, for a l l 0 ~ < j ~ m . In summary, at the beginning of the standard stage the right end marker of the r ight-most segment initializes m -[- 1 waves with rates z o, , z 1, ,..., z ,~ , , respectively, and the left end marker of the r ight-most segment initializes m -]- 1 waves with rates z0, z 1 .... , z ~ , respectively. For 0 ~ j ~ m, the wave from the left end marker with rate z~ meets the wave from the right end-marker with rate z~., at t ime wl at the point marked by j in Fig. 4. The meeting of these waves sets up the markers for the beginning of the next standard stage exactly in the way we desired to do it for our induction. The situation for the left-most segment is exactly analogous. There are + 1 waves from both the left end-marker and the right end-marker of the left-most segment. These move with rates z i and z ( where ( rw ) N = 2 - - - - / - - } - 1 , J z ~ - 2w ' j - - 2 z~, = 2w For all the intermediate segments we need only three waves init iated from each of the end-markers. F rom the left end-marker we need waves which propagate at rates z 0 , z 1 and z2 , from the right end-marker we need waves that propagate at rates z 0, , z 1, , and ze , . I t is quite easy to show (see Fig. 4) that after wl steps, the waves from the left end-marker which propagate at rates Zo, z l , and z e , respectively, meet the waves from the right end-marker which propagate at rates Zo, , z l , , and z 2, , respectively. The meeting points of two of these three pairs of waves are precisely the centers of the cells which were the left end-marker and right end-marker of the segment under consideration at the beginning. These again become markers for the beginning of the next stage. The meeting point of the third pair of waves is precisely the center of this segment, and it either falls into the center of a cell (if l is even), turning it into a marker, or on the boundary of two cells (if l is odd), turning both into markers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001652_j.jsv.2008.12.018-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001652_j.jsv.2008.12.018-Figure2-1.png", + "caption": "Fig. 2. Diagram of driveshaft in test bench.", + "texts": [ + " The non-constant velocity joint is located just to the front of the center bearing assembly. The second non-constant velocity joint is located at the rear axle end of the driveshaft. In addition, a vibration absorber, internal tuned damper (ITD), is incorporated in the front shaft to minimize vibrations due to the first bending mode of the driveshaft. The test bench is made up of a hydraulic motor that rotates the shaft on the transmission side, and a spindle at the opposite end to support the axle side of the driveshaft as shown in Fig. 2. Finally, a support for the center bearing is in the middle. This support was designed to adjust the height of the center bearing. This adjustment is used to test the driveshaft with different universal joint angles by changing the height of the center bearing. It is significant as the system has no brake or secondary motor to provide additional torque. Therefore, the \u2018\u2018load\u2019\u2019 torque in the shaft system is primarily due to inertia. This fixture has the capability of testing the driveshaft through a range of speeds" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure3-1.png", + "caption": "Fig. 3. Defection of cable cause by rigidity.", + "texts": [ + " Finally, the theoretical formula of the transmission capability including bending rigidity for precise cable drive is obtained in this section. The cablewill be bow-out in case the rigidity exists. Thus, the tangents to the cable at the ends of its arc of contactwith the pegwill be inclined to the lines of action of the forces pulling on the cable end. Thus the wrapped angular will be less than the total angular displacement of the cable between the points of application of the holding forces. The non-contact region of the cable can be treated as a one-end clamped beam,which is deflected by the tension forceT1 0 , shown in Fig. 3. On the condition of the bending rigidity, the inclined angle of load\u03c91 has been deduced byM.Wei and Rongjuan Chenwith the assumption of linear bending stiffness and small bending deflections [13], as where \u03c91 \u00bc 1 R1 ffiffiffiffiffi B T1 s \u00f07\u00de when \u03c91 \u2192 0, Q 0\u00f0 \u00de \u00bc T1 sin \u03c91 \u00bc 0 and T 0\u00f0 \u00de \u00bc T1 cos \u03c91 \u00bc T1 \u00f08\u00de , Q(0) and T(0) is the shear force and tension force at the end of the cable, respectively, which is also the boundary condition of ntact region. the co The wrapped angle will be decreased because of the bending rigidity of steel cable, as \u03b8e \u00bc \u03b8\u2212\u03b81\u2212\u03b82 \u00f09\u00de where \u03b81 \u2248 1 2 \u03c91 \u00bc 1 2R ffiffiffiffiffi B T1 s \u03b82 \u2248 1 2 \u03c92 \u00bc 1 2R ffiffiffiffiffi B T2 s 8>>< >>: \u00f010\u00de The effect of bending stiffness in the no-contact region can be concluded that the load inclines and the wrapped angle decreases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003449_imece2011-63452-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003449_imece2011-63452-Figure1-1.png", + "caption": "FIGURE 1: Model with flexible ring gear and defective planetary bearing.", + "texts": [ + " MODEL DESCRIPTION The model described here considers the in-plane dynamics of components. Sun, planets and carrier are modelled as rigid bodies (similar to Lin and Parker [6]) with coordinates q(\u2022) = (x(\u2022),y(\u2022),\u03b8(\u2022))T , where x(\u2022) and y(\u2022) are translational degreesof-freedom (DOF) in X and Y directions respectively, \u03b8(\u2022) is the rotational DOF, and subscript (\u2022) = s (for sun), pi (for ith planet) and c (for carrier). In addition to gears, bearing containing a localized defect is also included in the model (fig 1). Planet gear itself serves as the outer-race of the bearing and inner-race is modelled as a rigid body with coordinates qb = (xb,yb,\u03b8b) T . Inner and outer raceways are connected together with linear springs representing bearing stiffness kbp. Contact between meshing gears is represented by linear time-invariant springs. Ring gear is modelled as a continuous elastic ring. wr and w\u03b8 are the ring gear coordinates in radial and tangential directions. Equations of motion described in this section are derived in the XY Z coordinate system (CS) which is rotating with carrier", + " Therefore, Mcq\u0308c + { KB +Kcb 11 + i6=\u0393 \u2211 i=1:Z ( Kcp 11 ) i } qc +Kcb 12qb + i6=\u0393 \u2211 i=1:Z {( Kcp 12 ) i qpi } = 0 (16) and for sun, Msq\u0308s + { KB + Z \u2211 i=1 ( Ksp 11 ) i } qs + Z \u2211 i=1 {( Ksp 12 ) i qpi } = 0 (17) where Mc = diag(mc,mc, Ic), Ms = diag(ms,ms, Is) and KB = diag(kx b,k y b,0) is the stiffness matrix of bearings supporting sun and carrier. Mesh stiffness matrices are defined in the Appendix. Equation of Motion for Combined System The equation of motion of the combined system in the fixed CS (X fYf Z f fig. 1) can be written as Msys \u00a8\u0304q+ K\u0304sysq\u0304 = F\u0304 (18) where Msys is the mass matrix, K\u0304sys is the stiffness matrix in fixed CS, F\u0304 is the external force matrix and q\u0304 is the coordinate matrix in fixed CS. K\u0304sys will vary with time due to change in planet position caused by the rotation of the carrier. To avoid the time-varying stiffness in eq. 18 we can take advantage of the cyclic symmetry of the structure and formulate the equation of motion in the coordinate system XY Z (fig. 1) which is rotating with carrier. Now, system coordinate matrix (q) in rotating CS is q = Tq\u0304 or q\u0304 = TT q (19) where T is the transformation matrix (defined in Appendix). Substitution of eq. 19 into 18 yields Msysq\u0308+\u03c9cCLq\u0307+ ( Ksys\u2212\u03c92 c Kc ) q = F (20) with F = (Fs,Fr,Fp1 , \u00b7 \u00b7 \u00b7 ,FpZ ,Fb,Fc) T = TF\u0304, Ksys = TK\u0304sysTT , CL is a skew-symmetric Coriolis matrix, and Kc is the centripetal matrix. Carrier speed \u03c9c for a wind-turbine planetary drive is very small (typically 10-15 rpm), therefore, the effects of Coriolis and centripetal terms are neglected" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000135_s1560354708040096-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000135_s1560354708040096-Figure2-1.png", + "caption": "Figure 2", + "texts": [ + " In formula (2.2), \u039b = z \u2212 h(\u03b8), where h(\u03b8) is the distance between the point G and the support plane measured at contact. Therefore, \u039b\u0308 = z\u0308 \u2212 h\u2032\u2032(\u03b8)\u03b8\u03072 \u2212 h\u2032(\u03b8)\u03b8\u0308. (3.2) Taking (2.1) and (3.1) into account, we see that (3.2) becomes \u039b\u0308 = N m \u2212 f \u2032(\u03b8) A Ml \u2212 g \u2212 h\u2032\u2032(\u03b8)\u03b8\u03072 \u2212h\u2032(\u03b8) A \u03c8\u0307 sin \u03b8 ( (A \u2212 C)\u03c8\u0307 cos \u03b8 \u2212 C\u03d5\u0307 ) , (3.3) where Ml is the moment of inertia with respect to the straight line l that is parallel to the support plane and passes through the point G perpendicularly to the axis of the body (Fig. 2). In (3.3), the first two terms depend on the reaction of the plane and the other terms form the quantity U in (2.5). Therefore, U = \u2212g \u2212 h\u2032\u2032(\u03b8)\u03b8\u03072 \u2212 h\u2032(\u03b8) A \u03c4 (A\u03c4 cot \u03b8 \u2212 Cn) \u03c4 = \u03c8\u0307 sin \u03b8, n = \u03c8\u0307 cos \u03b8 + \u03d5\u0307. (3.4) REGULAR AND CHAOTIC DYNAMICS Vol. 13 No. 4 2008 As is known [3], a axisymmetric body on a plane can perform stationary motions for which the nutation angle has a constant value: \u03b8 = \u03b80 (the so-called regular precessions). Since for such motions, the total mechanical energy of the body is conserved, they are possible in the following two cases: either the plane is absolutely smooth or the body rolls on the plane without sliding (the cases of absolute roughness or the Amonton\u2013Coulomb law for the static friction)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure2-1.png", + "caption": "Fig. 2. (Color online) Configuration of cable drive system.", + "texts": [ + " For predigesting the behavior of the cable in precise cable drive mechanism, the bending stiffness B for the given cable with the samematerial, structure, radius and the samework condition is assumed to be a constant. Precise cable drive usually consisted of an output drum,with diameterDoutput, and an input drum,with diameterDinput, whichwas linked by cable segments. The precise cable is usually wrapped around the drive drum for several circles between the cable and the drum in a figure-eight pattern to increase the contact angle. The typical configuration of precise cable drive is shown in Fig. 2. A preload force must be applied to the cable initially to eliminate the transmission backlash and increase the stiffness of precise cable drive. When an external torque is applied to the drums, one part of the cable stretches while the other part slackens according to the direction of the torque. The tension difference between the incoming cable and outgoing cable balances the external torque [12,13], which can be determined as: MExt \u00bc T2\u2212T1\u00f0 \u00der1 \u00f03\u00de T1 \u00fe T2 \u00bc 2Tp \u00f04\u00de whereMExt, T1, T2, and r1 are the external torque, the outgoing tension, the incoming tension and the radius of input drum, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002852_s0022112010006075-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002852_s0022112010006075-Figure11-1.png", + "caption": "Figure 11. (a) Visualization of the helical tube inscribed in a torus. (b) VB/2\u03c0R versus s0 computed for the helical ring. In addition, the scaled translational velocities for the circumscribing torus predicted by the asymptotic (asym.) theory from Leshansky & Kenneth (2008) are shown, along with regularized Stokeslet computations of these toroidal swimming velocities.", + "texts": [ + " Would the twirling of each of the circular cross-sections about the circular torus axis, at constant angular velocity, cause the torus to swim in a Stokes fluid? Recently, Leshansky & Kenneth (2008) analysed this toroidal swimming due to surface rotations and showed that the swimmer is propelled against the direction of its outer surface. For the ratio s0 = b/a, where b is the radius of the circular axis of the torus, a is the cross-section radius, and us the surface speed of each cross-section, they compute a translational velocity V \u2248 (us/2s0)(log 8s0 \u2212 (1/2)) in the limit of large s0. Figure 11 allows us to visualize the relationship between our waving cylindrical ring and the toroidal swimmer, by inscribing the cylindrical ring in a torus whose radius is b = 0.5 and the cross-sectional radius is a = R + rh. For the torus, each point on a circular cross-section rotates with a constant angular velocity. On the other hand, only a segment of the helical tube, while tracing out roughly the same cross-section of the circumscribing torus, contributes an angular velocity. Figure 11 shows the comparison of the computed scaled translational velocity versus the ratio s0 for the helical ring, along with the asymptotic results for the circumscribing torus in Leshansky & Kenneth (2008). As expected, the translational velocity of the torus is greater than that of the inscribed helical ring. In addition, we computed the translational velocity of the toroidal swimmer using steady Stokeslet calculations, and these values show good agreement with the asymptotic theory (see figure 11). In an effort to understand the function of the longitudinal flagellum of dinoflagellates, we have formulated a simple fluid mechanical system of a waving cylindrical ring moving in a Stokes fluid. We conclude that the travelling wave imposed around the ring induces both rotational and translational motions. We have also used this model system to compare regularized Stokeslet formulations with slender-body theory, two popular numerical methods for Stokes flow. When computing rotational velocities of the ring, slender-body theory calculations agree very well with regularized Stokeslet calculations for thinner rings with small helix amplitudes and rings supporting smaller number of helical pitches" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000171_iros.2008.4650592-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000171_iros.2008.4650592-Figure4-1.png", + "caption": "Fig. 4. Definition of coordinate system.", + "texts": [ + " First, the trajectory space maps easily to the UGV actuation space (generally consisting of the throttle and steering angle). Navigation algorithms performed in the trajectory space will select command inputs that obey vehicle nonholonomic constraints. Second, dynamic constraints related to UGV rollover and side slip are easily expressible in the trajectory space, since these constraints are functions of the UGV velocity and path curvature. These constraints can also capture effects such as terrain inclination and roughness. The coordinate systems used in this work are shown in Fig. 4. A body frame B is fixed to the vehicle, with its origin at the vehicle center of mass. The position of the vehicle in the inertial frame I is expressed as the position of the origin of B. The vehicle attitude is expressed by x-y-z Euler angles using the vehicle yaw \u03b8, roll \u03c6, and pitch \u03c8 defined in B. To perform navigation a potential field is constructed in the trajectory space based on dynamic constraints, waypoint locations, and hazard locations. These issues are discussed below. Potential fields are defined based on UGV rollover and side slip constraints", + " If the predicted velocity profile exceeds the maximum safe velocity profile at any point, the robot\u2019s desired velocity is reduced to the maximum safe velocity along the trajectory. This approach attempts to exploit the computational efficiency of the potential field-based navigation with the safety guarantees implicit in optimal trajectory planning. Since low-order rigid body models are used in forward simulation, computational demands are negligible. V. SIMULATION RESULTS Dynamic simulations were conducted of a small UGV traveling at high speeds over uneven terrain. The UGV parameters were as follows (see Fig. 4): length L = 0.27 m, half-width d = 0.124 m, c.g. height h = 0.06 m, wheel diameter = 0.12 m. The potential field parameters were set as follows: Kr = 800, Ks = 800, Kg = 0.3, Kv1 = 0.5 x 10-5, Kv2 = 4, Ko = 1500, Kod = 0.05, Koa = 10, Kov = 0.07, T = 1.0 s. The gains were determined empirically, from simulation studies on a small number of canonical scenarios. Randomized rough terrain was generated using a fractal method modified to incorporate gross terrain undulation and discrete \u201cpeaks.\u201d Rough terrain with fractal number of 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001467_tcst.2008.2009528-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001467_tcst.2008.2009528-Figure1-1.png", + "caption": "Fig. 1. Experimental setup. (a) Measurement photo of a plane-type precision positioning table. (b) Experimental setup.", + "texts": [ + " In order to implement more degrees of freedom in a single plane, three directions are designed in one plane, where the -axis is attached to the -axis and the plane rotates about the -axis. In this paper, the positioning table is designed in series to avoid interference problem between axes, and the mass center and geometric center are required to locate at a same point as possible. By using three piezoelectric actuators to drive the plane-type 3-DOF precision positioning table, the input\u2013output relationship is easily obtained. Fig. 1 shows the top view of the plane-type 3-DOF precision positioning table assembled with three piezoelectric actuators, which have maximum displacement of 3 m. The table with material SKD 11 has the size of 60-mm length, 60-mm width, and 10-mm height. 1) Dynamic equations: In the differential equations, we only consider the cross-coupling effect in the stiffness matrix, and the general dynamic equation of the table is easily obtained as follows: (1) where the subscripts , , and represent the parameters in their corresponding directions, and are the equivalent masses, is the mass moment of inertia, , , and are the viscous damping coefficients, , , and are the equivalent spring stiffnesses, , , and denote the cross-coupling stiffnesses due to the topology of the stage and fabrication errors of the flexible mechanism in the -, -, and -axes, respectively, and is the rotation radius for the -axis", + " Finally, substituting (24a), (24b), and (24c) into (22) and designing the corresponding adaptation laws of the proposed ABC, we obtain (25) From (25), we can rearrange time derivative of Lyapunov stability function as follows: (26) In order to satisfy the Lyapunov stability theorem, we designate the adaptation laws as follows: (27) (28) (29) By substituting (27), (28), and (29) into (26), we can obtain the following result: (30) As a consequence, the control laws of (24a), (24b), and (24c) and the adaptation laws of (27), (28), and (29) for the ABC controller satisfy the Lyapunov stability theorem. Moreover, if all the values of control parameters in the -, -, and -axes are designed approximately, the asymptotical trajectory tracking for stability control and the improvement in the hysteresis behavior of the plane-type precision positioning table system can be completed. Fig. 1(a) shows the measurement photo of the plane-type 3-DOF precision positioning table. The measurement mass is to be measured when the table moves, and the capacitor-type gap sensor is a feedback measurement device of displacement signal. Fig. 1(b) shows the experimental setup for controlling the plane-type 3-DOF precision positioning table. The waveform of the applied input voltage is generated by the LabVIEW software, which is a Windows-supported graphical programming language. The D/A converter (NI PCI-6733E) with a resolution of 12 bit is used to transform the waveform to the power amplifier which has the voltage output range of 100 V, and then, send to the piezoelectric actuators of the table. The way to understand the displacements produced from piezoelectric actuators are by examining voltages from the capacitor-type gap sensor, transform to displacement signal converter and via A/D converter sensor amount (NI PCI-6052E), and analyze results from the response signal which we acquired" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001283_roman.2009.5326127-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001283_roman.2009.5326127-Figure5-1.png", + "caption": "Fig. 5. The position of the detected cluster and the interest region", + "texts": [ + " 4(a), and a new data is added which makes the cluster size bigger than the maximum size as Fig. 4(b), the number of clusters increases to make all sizes of clusters smaller than the maximum size like in Fig. 4(c). All humans or objects around a robot were detected by this method regardless of the number of them, and the variable number of the clusters allowed for addition and the deletion of clusters in the dynamic environment. The mean value of the data (r ) and the center data (\u03b8 ) in the detected cluster are compose the position (2) of the cluster as Fig 5. If many objects or humans are detected at every time step and they are moving, each human should be tracked with own id. In addition once the robot moves all the detected position of objects or humans change. In order to track the detected clusters in the flow of time step, the variable interest region was defined. In the Fig. 5 there are square shape interest region around the cluster, and the variables hF, hR, wL, wR are the size of the region. If there is a cluster in the previous cluster\u2019s interest region at the next time step, the clusters can be regarded as same human or object and it is marked as a same id. In addition, if the robot moves, the four variables can change in accordance with the linear and rotational velocities of the robot. The second problem was to obtain absolute human action (4) by estimating the absolute human velocity vector VH in (3) from the detected relative human position vector (2) in the first problem and measured robot velocity vector VR in (3)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002524_3.5535-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002524_3.5535-Figure1-1.png", + "caption": "Fig. 1 Euler angles for three-degree-of-freedom rotational motion.", + "texts": [ + " The present study treats the vehicle motion during the period in which, because of misalignment with the flight path during re-entry, the angle of attack is large relative to the trim angle of attack from mass or configurational asymmetries. The quasi-steady or mean value of the angle-ofattack oscillation envelope is obtained as a function of time (or altitude) during this period. II. Equations of Motion The vehicle motion is described in terms of the Euler angles \\l/, <\u00a3, 6, which describe the position of a set of bodyfixed axes x,y,z relative to nonrotating axes X,Y,Z that translate with the vehicle, as shown in Fig. 1. The axes \u2022f, 77, f are axes of roll, pitch, and yaw, respectively, relative to the plane of total angle of attack. They precess about the velocity vector with angular rate ^. If the principal moments of inertia about the \u00a3, 77, f axes are Ix, I, I, respectively, and the aerodynamic moments about these axes are MS, MT,, Mf, the moment equations of motion in terms of the Euler angles may be written8 MS = Ix(dp/dt) Mr, = 16 + Ixp\\l/ sin/9 - Ji/'2 sin0 cos0 sin0) + 10$ cos0 - where the roll rate p is defined by p = + $ cos0 (1) (2) The angular rate is the windward-meridian rotation rate, i", + " ^ These definitions are more restrictive than are the classical definitions of nutation and precession as being variations in 0 and if/, respectively (see, for example, Ref. 9, p. 432). D ow nl oa de d by N O R T H D A K O T A S T A T E U N IV E R SI T Y o n Ja nu ar y 25 , 2 01 5 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /3 .5 53 5 DECEMBER 1969 ROTATION RATE OF ROLLING RE-ENTRY VEHICLES 2327 The Euler angles \\f/, $, 6 were defined with respect to an inertial frame of reference that moves along the flight path such that \\l/ represents precession about the velocity vector (Fig. 1). Therefore, the two exoatmospheric conditions represented in Figs. 2c and 2d correspond to initial conditions on \\l/ of \\//Q = 0 and \\l/Q = 0 = jup0/cos0o, respectively. Referring back to the quasi-steady solution for ^, Eq. (8), we note that, in the limit as co -* 0, corresponding to zero atmospheric density, \\[/ can have the two values 0 and up (the latter value would be /xp/cos0, without the small angle approximations in Eq. 7). Therefore, the two precession modes of Eq. (8) correspond initially to the two cases of exoatmospheric motion shown in Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000704_s11783-009-0028-1-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000704_s11783-009-0028-1-Figure1-1.png", + "caption": "Fig. 1 Schematic (side view) (a) and picture (b) of the Quickscreen system", + "texts": [ + " The CEM (Ultrex CMI 7000, Membranesinternational, USA) was soaked in 3% NaCl for at least 24 h before being used in the MEA. All components of the MEA were cleaned thoroughly in deionized water. First, the carbon paper electrodes were pressed onto the CEM, forming an anode-membranecathode assembly. Next, it was cleaned by successively boiling in 3% hydrogen peroxide, 0.5 mol/L sulfuric acid, and deionized water, each for 1 h. 2.3 Construction and operation of the Quickscreen system The system was assembled as shown in Fig. 1(a), where the main body of the Quickscreen system comprised three polycarbonate plates and one MEA, with a size of 7 cm by 7 cm and the height of 0.6 cm. The MEA was placed between the lower two plates to separate the anode and cathode chambers. The cross-section of the working area of the anode was about 0.8 cm2 and the anode chamber volume was about 200 \u03bcL. In order to prevent oxygen diffusion during the experiment, the anode chamber was sealed with a thick rubber stopper held in place by the upper polycarbonate plate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000446_08ias.2008.31-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000446_08ias.2008.31-Figure3-1.png", + "caption": "Figure 3. Simplified model of the mechanical part using damping and stiffness coefficients.", + "texts": [ + " In the considered electromechanical system, there are also other elements which are not taken into account due to the small transmission shaft length between the induction machine and the pinion and between the wheel and the mechanical load (Fig. 2). These last mechanical elements can be considered as rigid connections. This assumption gives the opportunity to integrate the wheel inertia with the load one and the pinion inertia with the one of the induction machine rotating part in order to simplify the mechanical model (Fig. 3). Considering a zero backlash transmission system, the gearbox dynamic model can be written as [18]: ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) 1 1 1 1 1 2 2 1 1 1 2 2 2 2 2 1 1 2 2 2 1 1 2 2 M z e L z L J J t r K t r t r t e t r d r t r t e t T t J J t r K t r t r t e t r d r t r t e t T t \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u23a1 \u23a4+ + + +\u23a3 \u23a6 \u23a1 \u23a4+ + + =\u23a3 \u23a6 \u23a1 \u23a4+ + + +\u23a3 \u23a6 \u23a1 \u23a4+ + + =\u23a3 \u23a6 (15) with, J1 : pinion inertia J2 : wheel inertia JM : induction machine rotor inertia JL : load inertia r1, r2 : radius of pinion and wheel dz : damping coefficient of contact point K(t) : stiffness function of contact point e(t) : transmission error Te(t) : electromagnetic torque TL(t) : load torque \u03b81(t) : rotational angle of the pinion \u03b82(t) : rotational angle of the wheel Expression (15) can be simplified using a new coordinate system represented by (q1, q2), with the following relations: ( )( ) ( ) ( ) ( )1 2 1 1 2 1 2 M L e LJ J J J T t T t q t m rr r r \u239b \u239e+ + = \u2212\u239c \u239f\u239c \u239f \u239d \u23a0 (16) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) 1 2 2 2 2 1 2 ( ) e L z M L z rT t r T t m q t d q t K t q t m J J J J K t e t d e t \u239b \u239e + + = +\u239c \u239f\u239c \u239f+ +\u239d \u23a0 \u2212 \u2212 (17) where q1 , q2 are defined as: ( ) ( ) ( )2 1 1 1 2 2 1L M r r q t m t t J J J J \u03b8 \u03b8 \u239b \u239e = \u2212\u239c \u239f+ +\u239d \u23a0 (18) ( ) ( ) ( )2 1 1 2 2q t r t r t\u03b8 \u03b8= + (19) with: ( )( ) ( ) ( ) 1 2 2 2 1 2 2 1 M L M L J J J J m J J r J J r + + = + + + (20) Therefore, two completely decoupled expressions can be used for the mechanical system model and for simulation purpose" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002325_s10846-011-9584-2-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002325_s10846-011-9584-2-Figure1-1.png", + "caption": "Fig. 1 C-Plane II UAV", + "texts": [ + " 2 The C-Plane II Vehicle This section addresses the operational description and the equations of motion of the aerial vehicle (C-Plane II UAV). Aerodynamic and wind effects of the longitudinal dynamics of this aircraft are studied, for this, the nonlinear mathematical equations to represent this system are derived using the Euler\u2013Lagrange formulation. 2.1 Operational Description The C-Plane II vehicle is a tail-sitter VTOL UAV with a propulsion system having a pair of brushless motors in coaxial configuration (see Fig. 1). This vehicle is capable of performing vertical flight as a helicopter and operating horizontal flight with the same effectiveness as a conventional airplane. Moving through the air, the vehicle utilizes the lift produced by outer body lift surfaces in order to attain the horizontal flight. For exploring the forward flight stability, the mini aircraft possesses aerodynamic control surfaces that allow rotational degrees of freedom in order to control pitch, roll and yaw motion. The aileron-elevon system controls the pitch and roll coupled motion, and the differential speed control of the two rotors also regulates the roll motion whereas the direction of the vehicle is manipulated by the rudder" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000076_j.robot.2008.01.002-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000076_j.robot.2008.01.002-Figure2-1.png", + "caption": "Fig. 2. Three-joint leg model.", + "texts": [ + " If \u03b1 = 0, the quadruped robot will have straight-line walking where the center of gravity traces on the X -axis and each leg tracks on the middle line of its working area (see Fig. 1), paralleled with the trajectory of the center of gravity. \u03b1d , a robot parameter that will be used later in this paper, is defined as the angle between the off-diagonal and the base of the working area. It is supposed that a leg attached to the quadruped robot has the geometry of the articulated arm [8] shown in Fig. 2. This model has two rigid links and three revolute joints; the lower link is connected to the upper link via an active revolute joint and the upper link is connected to the body via two active revolute joints, one parallel with the knee joint and the other parallel with the body longitudinal axis. Hence the foot point has three degrees of freedom with respect to the body and the overall walking can be driven in any direction. We denote the joint at the main actuator as joint one, the joint at the lifting actuator as joint two and the joint at the knee actuator as joint three", + " Hence a locked joint failure happening to a leg with three degrees of freedom will result in twodimensional motions of the leg and thus will forbid the failed leg to have normal swing in the transfer phase or backward movement in the support phase. Nevertheless, unlike free-swinging failure [12] and mutilation failure, a locked joint failure does not take away body-supporting ability from the failed leg. For employing the failed leg in post-failure walking, we should examine the configuration of the failed leg determined by the position of a locked joint and the resulting change of the reachable area. Fig. 3 illustrates the behavior of a failed leg with the geometry of Fig. 2. After joint one of a leg is locked from failure, the kinematics of the failed leg is the same as a two-link revolute joint manipulator. Its workspace is reduced to the plane made of the two links and the reachable region of the foothold position in the working area is projected onto a straight-line of which the lateral view is shown in Fig. 3(a). \u03b8\u03021 denotes the locked angle of joint one, and the values of \u03b82 and \u03b83 are determined by the foothold position. The locked failure at joint two or joint three yields almost identical post-failure behavior" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.8-1.png", + "caption": "Fig. 2.8 The model of an Anchor escapement", + "texts": [ + " Therefore, as new and better designs emerged, the verge escapement gradually disappeared in 1800s. The Anchor escapement is another milestone invention. It was perhaps invented by the famous British scientist Robert Hooke (1635\u20131703) around 1657, as shown in Fig. 2.7 (Wikipedia 2002a), and first made by the British clock master Thomas Tompion (1639\u20131713). However, like many of his other works, his ownership is disputed (Wikipedia 2004c). In any case, Hooke\u2019s milestone contribution to mechanical watch and clock making is indisputable. Figure 2.8 shows the model of the Anchor escapement. It consists of an escape wheel, an anchor and a pendulum. The exact shapes of both the escape tooth and the anchor pallet are not crucial. The escape wheel is driven by a lifted weight or a wound mainspring rotating clockwise. As a tooth of the escape wheel slides on the surface of the left pallet of the anchor, the anchor moves away releasing the tooth and allowing the escape wheel to advance. Next, the pendulum reaches its highest position and swings back" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001846_iccas.2010.5670292-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001846_iccas.2010.5670292-Figure1-1.png", + "caption": "Fig. 1 The homing guidance geometry.", + "texts": [ + " The remainder of this paper will discuss number of simulation studies to demonstrate and evaluate the performance of proposed guidance law. To determine the improving capability of target observability, the modified gain pseudo-measurement filter (MGPMF)[6] is used for the estimator because of its good estimation performance as demonstrated in the pervious work [6, 7]. Finally, conclusions are offered in section 5. In this section, the engagement kinematics is described. A considering homing guidance geometry for a stationary and a slowly moving target is depicted in Fig. 1, where (XI , YI) denotes a Cartesian inertial reference frame and (Xf , Yf ) denotes an impact angle frame. The subscripts M and T represents the missile and target, respectively. The missile flight path angle, speed and normal acceleration are expressed by \u03b3M , VM and aM , respectively. R is the relative distance and \u03bb is the Lineof-sight (LOS) angle between a target. \u03c8e represents the heading error. Before deriving kinematics equations, it is assumed that a homing guidance law follows some accurate guidance handover from a proper mid-course guidance logic so that a lateral miss distance with respect to the impact angle frame and a missile heading error are small at the initial homing phase", + " Firstly, the simulation for showing the ability of impact angel control is performed. And then the improving capability of target observability employing the proposed guidance law is proven by comparing other guidance law. Especially, the conventional proportional navigation guidance (PNG) law [8] is adopted. The modified gain pseudo-measurement filter (MGPMF) as covered in Ref. [6] is used for the estimator and the singer\u2019s work [9] is chosen for the target model. The homing geometry as described in Fig. 1 with following parameters are used to setup the test engagement scenario. Firstly, we outline parameters for homing geometry. The initial relative distance is R0 = 3km. A stationary target is placed on (8000, 0)m. The constant missile speed is assumed to be VM = 200m/s. The initial relative distance with respect to the impact angle frame is y0 = 100m. The initial heading error and desired impact angle are chosen as \u03c8e = 10\u25e6 and \u03b3f = 45\u25e6, respectively. To take account of noisy measurements, two types of accelerometer error sources such as velocity random walk and bias stability are considered" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001213_s1874-1029(08)60091-9-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001213_s1874-1029(08)60091-9-Figure4-1.png", + "caption": "Fig. 4 The state variables xv trajectories", + "texts": [], + "surrounding_texts": [ + "No. 6 WEI Qing-Lai et al.: Data-based Optimal Control for Discrete-time Zero-sum \u00b7 \u00b7 \u00b7 689\nStep 1. Give the boundary condition xh(0, l) = f (l) and xv(k, 0) = g(k). Let P0 = 0, K0 = 0, and L0 = 0. Give the computation accuracy \u03b5.\nStep 2. According to the N sampling points, compute\nZN and Y\u0302 N according to (83) and (84). Step 3. Compute hi according to (82) and Hi according to (84) through the Kronecker method. Step 4. Compute the feedback control laws by\nKi+1 = (Hi uu \u2212 Hi uw(Hi ww)\u22121Hi wu)\u22121\u00d7 (Hi uw(Hi ww)\u22121Hi wx \u2212 Hi ux) (87)\nand\nLi+1 = (Hi ww \u2212 Hi wu(Hi uu)\u22121Hi uw)\u22121\u00d7 (Hi wu(Hi uu)\u22121Hi ux \u2212 Hi wx) (88)\nStep 5. If\n\u2016 hi+1 \u2212 hi \u2016\u2264 \u03b5 (89)\nexit; otherwise, go to Step 6. Step 6. Set i = i + 1, go to Step 2.\n4 Neural network implementation\nIn this subsection, neural networks are constructed to implement the iterative ACD algorithm. There are several ACD structures that can be chosen[27]. As HDP structure is basic and convenient to realize, we will use it to implement the iterative ACD algorithm.\nAssume the number of hidden layer neurons is denoted by l, the weight matrix between the input layer and hidden layer is denoted by V , and the weight matrix between the hidden layer and output layer is denoted by W . Then, the output of three-layer neural network is represented by\nF\u0302 (X, V, W ) = WT\u03c3(V TX ) (90)\nwhere \u03c3(V TX ) \u2208 Rl, [\u03c3(z)]i = ez\ni \u2212 e\u2212zi ez i + e\u2212zi , i = 1, \u00b7 \u00b7 \u00b7 , l, are\nthe activation function. The neural network estimation error can be expressed by\nF (X ) = F (X, V \u2217, W \u2217) + \u03b5(X ) (91)\nwhere V \u2217 and W \u2217 are the ideal weight parameters, and \u03b5(X ) is the reconstruction error.\nHere, there are three neural networks, which are critic network, action network u, and action network w. All the neural networks are chosen as three-layer feedforward networks. The whole structure diagram is shown in Fig. 1. The utility term in the figure denotes xT(k, l)Qx(k, l) + uT(k, l)Ru(k, l) + wT(k, l)Sw(k, l).\n4.1 The critic network\nThe critic network is used to approximate the Hamilton function H(k, l). The output of the critic network is denoted as\nH\u0302i(k, l) = WT ci\u03c3(V T ci x(k, l)) (92)\nThe target function can be written as\nHi+1(k, l) = H+ i (k, l)+\n[ xT(k, l) uT i (k, l) wT i (k, l) ]\u23a1\u23a3Q 0 0 0 R 0 0 0 S \u23a4 \u23a6 \u23a1 \u23a3 x(k, l) ui(k, l) wi(k, l) \u23a4 \u23a6\n(93)\nThen, we define the error function for the critic network as\neci(k, l) = H\u0302i+1(k, l) \u2212 Hi+1(k, l) (94)\nAnd, the objective function to be minimized in the critic network is\nEci(k, l) = 1\n2 e2 ci(k, l) (95)\nSo the gradient-based weight updating rule[39] for the critic network is given by\nwc(i+1)(k, l) = wci(k, l) + \u0394wci(k, l) (96)\n\u0394wci(k, l) = \u03b1c [ \u2212\u2202Eci(k, l)\n\u2202wci(k, l)\n] (97)\n\u2202Eci(k, l) \u2202wci(k, l) = \u2202Eci(k, l)\n\u2202H\u0302i(k, l)\n\u2202H\u0302i(k, l) \u2202wci(k, l) (98)\nwhere \u03b1c > 0 is the learning rate of critic network and wc(k, l) is the weight vector in the critic network.\n4.2 The action networks\nAction networks are used to approximate the iterative optimal controls. There are two action networks, which are used to approximate the optimal controls u and w, respectively.\nFor the action network that approximates the control u(k, l), state x(k, l) is used as the input to create the optimal control and u(k, l) is used as the output of the network. The output can be formulated as\nu\u0302i(k, l) = WT ai\u03c3(V T aix(k, l)) (99)\nSo, we can define the output error of the action network as\neai(k, l) = u\u0302i(k, l) \u2212 ui(k, l) (100)\nwhere ui(k, l) is the target function that can be described by\nui(k, l) = (Hi ww \u2212 Hi wu(Hi uu)\u22121Hi uw)\u22121\u00d7 (Hi wu(Hi uu)\u22121Hi ux \u2212 Hi wx)x(k, l) (101)\nwhere Hi can be obtained according to Kronecker product in (85).\nThe weighs in the action network are updated to minimize the following performance error measure:\nEai(k, l) = 1\n2 e2 ai (102)", + "690 ACTA AUTOMATICA SINICA Vol. 35\nThe weight updating algorithm is similar to the one for the critic network. By the gradient descent rule, we can obtain\nwa(i+1)(k, l) = wai(k, l) + \u0394wai(k, l) (103)\n\u0394wai(k, l) = \u03b2a [ \u2212\u2202Eai(k, l)\n\u2202wai(k, l)\n] (104)\n\u2202Eai(k, l) \u2202wai(k, l) = \u2202Eai(k, l) \u2202eai(k, l) \u2202eai(k, l) \u2202ui(k, l) \u2202ui(k, l) \u2202wai(k, l) (105)\nwhere \u03b2a > 0 is the learning rate of the action network. For the action network w that approximates the control w(k, l), state x(k, l) is used as the input to create the optimal control and w(k, l) is used as the output of the network. The target of w action network can be expressed as\nwi(k, l) = (Hi ww \u2212 Hi wu(Hi uu)\u22121Hi uw)\u22121\u00d7 (Hi wu(Hi uu)\u22121Hi ux \u2212 Hi wx)x(k, l) (106)\nAll the update rules of w action network are the same as the update rules of u network and it is omitted here.\n5 Simulation\nIn this section, the proposed method is applied to an air drying process control. Our example is a modification of Example 1 in [40] and extends the variable space to the infinite horizon.\nThe dynamical processes can be described by the following Darboux equation:\n\u22022x(s, t)\n\u2202s\u2202t = a1\n\u2202x(s, t)\n\u2202t + a2\n\u2202x(s, t)\n\u2202x + a0x(s, t) +\nbu(s, t) + cw(s, t) (107)\nwith the initial and boundary conditions\nxh(0, t) = { 0.5, t \u2264 4 0, t > 4 , xv(s, 0) = { 1, s \u2264 4 0, s > 4 (108)\nwhere x(s, t) is an unknown function, a0, a1, a2, b, and c are real coefficients, u(s, t) and w(s, t) are the input functions. The variable x means the humidity, which is the system state, s means the location of the air, and t is the processing time.\nLet a0 = 0.2, a1 = 0.3, a2 = 0.1, b = 0.3, and c = 0.25. The quadratic performance index function is formulated as\nV =\n\u222b \u221e\nt=0\n\u222b \u221e\ns=0\n{ Qx2(s, t)+Ru2(s, t)+Sw2(s, t) } dsdt\n(109)\nThe discretization method for system (107) is similar to the method in [40]. Suppose that the sampling periods of the digital control system are chosen as X = 0.1 cm and T = 0.1 s. Following the methodology presented in [40], we can compute the discretized system equation (1) as[\nxh((k + 1)X, lT ) xv(kX, (l + 1)T )\n] = [ 0.7408 0.2765 0.0952 0.9048 ] [ xh(kX, lT ) xv(kX, lT ) ] +[\n0.0259 0\n] \u00d7 u(kX, lT ) + [ 0\n0.0564\n] w(kX, lT )\n(110)\nwith the boundary conditions\nxh(0, lT ) = { 0.5, l \u2264 40 0, l > 40 , xv(kX, 0) = { 1, k \u2264 40 0, k > 40\n(111)\nand the discredited performance index function as V = \u221e\u2211\nk=0 \u221e\u2211 l=0 { Qx2(Xk, T l) + Ru2(Xk, T l) + Sw2(Xk, T l) } (112)\nWe implement the iterative algorithm at (k, l) = (0, 0). We choose three-layer neural networks as the critic network, the action network u, the action network w with the structures 2-8-1, 2-8-1, and 2-8-1, respectively. The initial weights of action networks and critic network are all set to be random in [\u22120.5, 0.5]. Then, the critic network and the action network are trained for i = 50 times so that the given accuracy \u03b5 = 10\u22126 is reached. In the training process, the learning rate \u03b2a = \u03b1c = 0.05. The evaluating point number N = 40 for every iteration and choose the small white noise as \u03be1(0, 0.01) and \u03be2(0, 0.01). The convergence curve of the performance index function is shown in Fig. 2. Then, we apply the optimal control to the system for k = 40, l = 40 time steps and obtain the following results. The state trajectories are given as Figs. 3 and 4. The control curves are given as Figs. 5 and 6, respectively.", + "No. 6 WEI Qing-Lai et al.: Data-based Optimal Control for Discrete-time Zero-sum \u00b7 \u00b7 \u00b7 691\nFrom the simulation results, we can see that the proposed iterative ACD algorithm in this paper obtains good effects. In [40], Tsai just studied the model-based optimal control in the finite horizon. In this paper, using the iterative ACD algorithm, the optimal control scheme for 2-D system in the infinite horizon can also be obtained without the system model. So the proposed algorithm in this paper is more effective than the method in [40] for industry process control.\n6 Conclusion\nIn this paper, we proposed an effective iterative algorithm to find the optimal controller of a class of discretetime two-person zero-sum games for Roesser types 2-D systems. The proposed ACD algorithm allows to be implemented without the system model. Stability analysis of the 2-D systems was presented and the convergence property of the performance index function was also proved. The simulation study has successfully demonstrated the upstanding performance of the proposed optimal control scheme for the 2-D systems.\nReferences\n1 Jamshidi M. Large Scale Systems: Modeling, Control, and Fuzzy Logics. Amsterdam: The Netherlands Press, 1982\n2 Chang H S, Marcus S I. Two-person zero-sum Markov games: receding horizon approach. IEEE Transactions on Automatic Control, 2003, 48(11): 1951\u22121961\n3 Chen B S, Tseng C S, Uang H J. Fuzzy differential games for nonlinear stochastic systems: suboptimal approach. IEEE Transactions on Fuzzy Systems, 2002, 10(2): 222\u2212233\n4 Nian Xiao-Hong, Cao Li. Design of optimal observer and optimal feedback controller based on differential game theory. Acta Automatica Sinica, 2006, 32(5): 807\u2212812 (in Chinese)\n5 Nian Xiao-Hong. Suboptimal strategies of linear quadratic closed-loop differential games: a BMI approach. Acta Automatica Sinica, 2005, 31(2): 216\u2212222\n6 Bertsekas D P. Convex Analysis and Optimization. Boston: Athena Scientific, 2003\n7 Goebel R. Convexity in zero-sum differential games. SIAM Journal of Control and Optimization, 2001, 40(5): 1491\u22121504\n8 Altman E, Basar T. Multiuser rate-based flow control. IEEE Transactions on Communications, 1998, 46(7): 940\u2212949\n9 Basar T, Olsder G J. Dynamic Noncooperative Game Theory. New York: Academic Press, 1982\n10 Basar T, Bernhard P. H\u221e Optimal Control and Related Minimax Design Problems. Boston: Birkhauser Press, 1995\n11 Hua X, Mizukami K. Linear-quadratic zero-sum differential games for generalized state space systems. IEEE Transactions on Automatic Control, 1994, 39(1): 143\u2212147\n12 Wei G, Feng G, Wang Z. Robust H\u221e control for discretetime fuzzy systems with infinite-distributed delays. IEEE Transactions on Fuzzy Systems, 2009, 17(1): 224\u2212232\n13 Werbos P J. Approximate dynamic programming for realtime control and neural modeling. Handbook of Intelligent Control: Neural, Fuzzy, and Adaptive Approaches. New York: Van Nostrand Reinhold, 1992\n14 Xu Jian-Ming, Yu Li. H\u221e control for 2-D discrete state delayed systems in the second FM model. Acta Automatica Sinica, 2008, 34(7): 809\u2212813\n15 Uetake Y. Optimal smoothing for noncausal 2-D systems based on a descriptor model. IEEE Transactions on Automatic Control, 1992, 37(11): 1840\u22121845\n16 Owens D H, Amann N, Rogers E, French M. Analysis of linear iterative learning control schemes \u2014 a 2D systems/repetitive processes approach. Multidimensional Systems and Signal Processing, 2000, 11(1-2): 125\u2212177\n17 Sulikowski B, Galkowski K, Rogers E, Owens D H. Output feedback control of discrete linear repetitive processes. Automatica, 2004, 40(12): 2167\u22122173\n18 Li C J, Fadali M S. Optimal control of 2-D systems. IEEE Transactions on Automatic Control, 1991, 36(2): 223\u2212228\n19 Liu D R, Javaherian H, Kovalenko O, Huang T. Adaptive critic learning techniques for engine torque and air-fuel ratio control. IEEE Transactions on Systems, Man, and Cybernetics, Part B: Cybernetics, 2008, 38(4): 988\u2212993\n20 Al-Tamimi A, Abu-Khalaf M, Lewis F L. Adaptive critic designs for discrete-time zero-sum games with application to H\u221e control. IEEE Transactions on Systems, Man, and Cybernetics, Part B: Cybernetics, 2007, 37(1): 240\u2212247\n21 Liu De-Rong. Approximate dynamic programming for selflearning control. Acta Automatica Sinica, 2005, 31(1): 13\u221218\n22 Ray S, Venayagamoorthy G K, Chaudhuri B, Majumder R. Comparison of adaptive critic-based and classical wide-area controllers for power systems. IEEE Transactions on Systems, Man, and Cybernetics, Part B: Cybernetics, 2008, 38(4): 1002\u22121007\n23 Watkins C. Learning from Delayed Rewards [Ph. D. dissertation], Cambridge University, USA, 1989\n24 Werbos P J. A menu of designs for reinforcement learning over time. Neural Networks for Control. Cambridge: MIT Press, 1991. 67\u221295\n25 Widrow B, Gupta N K, Maitra S. Punish/reward: learning with a critic in adaptive threshold systems. IEEE Transactions on Systems, Man, Cybernetics, 1973, 3(5): 455\u2212465\n26 Prokhorov D V, Wunsch D C. Adaptive critic designs. IEEE Transactions on Neural Networks, 1997, 8(5): 997\u22121007\n27 Murray J J, Cox C J, Lendaris G G, Saeks R. Adaptive dynamic programming. IEEE Transactions on Systems, Man, and Cybernetics, Part C: Applications and Reviews, 2002, 32(2): 140\u2212153\n28 Zhang H G, Wei Q L, Luo Y H. A novel infinite-time optimal tracking control scheme for a class of discrete-time nonlinear systems via the greedy HDP iteration algorithm. IEEE Transactions on Systems, Man, Cybernetics, Part B: Cybernetics, 2008, 38(4): 937\u2212942" + ] + }, + { + "image_filename": "designv11_3_0000002_iros.2007.4399084-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000002_iros.2007.4399084-Figure1-1.png", + "caption": "Fig. 1. Conceptual diagram of net structure tactile sensor", + "texts": [ + " Also, since every detection element needs scanning and processing sequence, response speed decreases. In this study, a net-structure tactile sensor attachable to freeform surfaces and with reduced wiring requirement was developed. The sensor features high response speeds without requiring software processing, is formed into a structure resembling a net, can be attached to freeform surfaces, and can detect the center of the load distribution as well as the overall load on 2-dimensional surfaces. Fig.1 shows the conceptual schematic diagram of the sensor. There are only 4 signals from the sensor and Internal connection is required only between adjacent detection elements. Since the structure is like that of a net, the system of sensors can cover 3-dimensional freeform surfaces. Sensor response speed is almost unaffected by the number, placement, or sensor surface area of the detection elements. In summary, the sensor features the following: 1) Covering freeform surfaces: the sensor can be laid out like a net to cover freeform surfaces 2) Reduced wiring: Internal connection is made between adjacent detection elements only" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003296_j.talanta.2011.09.013-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003296_j.talanta.2011.09.013-Figure1-1.png", + "caption": "Fig. 1. Flow-injection manifold configuration. S: sample or standard solution (ascorbic acid), CS: carrier stream (ultra pure water), RS: reagent stream (1.6 \u00d7 10\u22123 M iron(III) a \u22124 lve (lo D puter", + "texts": [ + " All spectral measurements and real-time data acquisition of flow injection peaks were obtained using a double beam UV/vis spectrophotometer (Shimadzu UV-1601, Kyoto, Japan) fitted with a flow through cell. The instrument was interfaced to a computer equipped with UV Probe 2.31 software provided by Shimadzu. Adjustments and measurements of pH were carried out using a Mettler Toledo SevenMulti potentiometer (Mettler Toledo, Schwerzenbach, Switzerland) equipped with combined glass electrode Mettler Toledo InLab\u00ae413. In the optimization part of experiment a temperature-controllable water bath accurate to \u00b10.5 \u25e6C was used. 2.3. Flow injection procedure In the developed flow system, depicted in Fig. 1, the loop (500 L) of the rotary valve was filled with the sample (or standard solution) while the ultra pure water carrier stream (CS) was mixed with the reagent stream (RS). RS consisted of 1.6 \u00d7 10\u22123 M iron(III) and 8.0 \u00d7 10\u22124 M TPTZ in acetate buffer, pH 3.6. CS and RS yielded the final stream that allowed the establishment of the baseline. By valve switching, the sample or standard solutions were injected in the carrier stream, and thus formed sample zone flowed to the confluence point (CP) where it was mixed with reagent stream" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002164_s12206-012-0865-x-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002164_s12206-012-0865-x-Figure3-1.png", + "caption": "Fig. 3. Schematic diagram of the flow geometry and the computational domain for the case of multiple particles submerged in an enough large square cavity (L = 40) subjected to a uniform external electric field. Note that all the lengths and spatial coordinates are nondimensionalized by the particle radius.", + "texts": [ + " In addition, 1k p +u is the velocity field on the particle domain computed from the force and torque balances (18) and (19) during the iteration. The flow chart for direct simulations on the DEP motions is plotted in Fig. 2, together with that on the MP motions, indicating the analogy of the simulation procedure between them. Consider that all the cylindrical particles are initially equispaced on a circle with an origin at the center of a square cavity (computational domain), that is, at , , ,( , ) ( cos ,c i c i c c ix y r \u03b8= ,sin ),c c ir \u03b8 where , 0 ( 1)c i c ci\u03b8 \u03b8 \u03b8= + \u2212 \u2206 ( 360 /c pN\u03b8\u2206 = \u00b0 and 00 360 /c pN\u03b8\u00b0 \u2264 \u2264 \u00b0 ). Fig. 3 shows the schematic diagram of flow geometry and computational domain, where a uniform electric field is externally applied in the horizontal (x) direction and the square cavity is set too large to affect the fluid flow around the particles. Along the cavity wall, therefore, zero velocity ( 0)=u and linearly-varying potential ( )x\u03d5 = \u2212 are applied respectively for the fluid flow and the electric potential as boundary conditions. Referring to Table 1, the DEP and MP motions are governed by the same or very similar nondimensional equations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002339_iros.2011.6095049-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002339_iros.2011.6095049-Figure2-1.png", + "caption": "Fig. 2. Sub-mechanism and the kinematic model of PA, with the end effectors set as reference coordinates (x = 0, y = 0, Z = \u00b0 at end effector).", + "texts": [ + " In this section, the mechanical structure and control algorithm for each module are briefly introduced. A pair of PA provides unrestricted pelvis motion using three translational movements to provide workspace for rotational movement of pelvis during walking. Ll2 xh = L+ \ufffd \ufffd 1/+l22 1112 Yh = ---Jl12+l22 (1) (2) where L = 20cm . The range of motion of the actuator is -Scm < l1 < Scm, Scm < l2 < lOcm, and -4cm < 53 < 4cm. Where Bx = 0; By = 0; Bz = 0, with the end effectors set as reference coordinates x 0= 0, Y 0= 0, Z 0= \u00b0 at end effector. The kinematic model is shown in Fig 2. The trajectory of pelvis is provided as A pair of robotic orthosis are connected to PA and parallel attached to lower limbs of subject. The hip, knee and ankle joints of subject can obtain the active assisted, while the metatarsal joints are passive spring connection without control. xm = Xh -l3sin(Bhip) - l4sin(Bknee -Bhip) + lscos(Bknee -Bhip+Bankle) Ym = Yh -l3cos(Bhip) - l4COs(Bknee -Bhip) + lssin(Bknee -Bhip+Bankle) (4) (5) where l3 = 40 em, l4 = 40 em and ls = 15 em, -450 < Bhip < 450, 00 < Bknee < 600, and -300 < Bankle < 300 , the kinematic model is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002790_s11465-013-0254-x-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002790_s11465-013-0254-x-Figure2-1.png", + "caption": "Fig. 2 Nonlinear dynamic model of the cam mechanism with oscillating roller follower", + "texts": [ + " This mechanism consists of a cam (1) mounted on a camshaft, an oscillating follower (3) with 2 rollers (2) and (4) and a sliding rod (5) that translate vertically. A spring is inserted between the sliding rod (5) and the frame to maintain two contacts at points C1 and C2. Cam (1) rotates at a constant angular velocity \u03c91. This rotation causes the oscillation of the follower (3) relatively to a fixed point O3 and hence the translation of the sliding rod (5). We suppose that we have always two contact points C1 and C2 between the roller (2) and Cam (1) and between the roller (4) and the sliding rod (5) respectively. Figure 2 represents the nonlinear dynamic model of the Received October 17, 2012; accepted March 7, 2013 Walha LASSAAD (\u2709), Tounsi MOHAMED, Driss YASSINE, Chaari FAKHER, Fakhfakh TAHER, Haddar MOHAMED Research Unit of Mechanical Dynamic System (UDSM), Mechanical Engineering Department, National Engineers School of Sfax, University of Sfax, Tunisia E-mail: walhalassaad@yahoo.fr cam mechanism with oscillating roller follower. The oscillating follower is considered as a rigid body mounted on a rigid bearing connected to the frame [6,7]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000892_s0263574709005748-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000892_s0263574709005748-Figure2-1.png", + "caption": "Fig. 2. n-dof manipulator operating in the two dimensional space.", + "texts": [ + " Thus, the reduction of the search space complexity has a great impact on the final optimum solution. In our approach, this is achieved by using the B-Surface concept to formulate a search space represented by a single mathematical entity. On the other hand, research on combinatorial explosion based on metaheuristics, such as GAs, can lead to approximate solutions in polynomial time instead of exact solutions that would be at intolerably high cost. The kinematics equations of a nth DOF planar manipulator with rotational joints operating in a 2D Cartesian Space (Fig. 2) is expressed by Eqs. (1)\u2013(2): xi+1 = x0 + i\u2211 k=0 k cos \u239b \u239d k\u2211 j=0 \u03b8j \u239e \u23a0, i = 0, 1, . . . , n \u2212 1, (1) yi+1 = y0 + i\u2211 k=0 k sin \u239b \u239d k\u2211 j=0 \u03b8j \u239e \u23a0, i = 0, 1, . . . , n \u2212 1, (2) where = ( 0, 1, . . . , n\u22121) \u2208 n are the lengths of the arm\u2019s links, \u03b8 = (\u03b80, \u03b81, . . . , \u03b8n\u22121) \u2208 n are the angle variables of the arm\u2019s joints and p = [(x0, y0), (x1, y1), . . . , (xk, yk), . . . , (xn, yn)] \u2208 (n+1)\u00d72 are the Cartesian coordinates of the arm\u2019s joints with respect to a fixed coordinate system. Inverse kinematics concerns the solution of the above kinematic equations, which are nonlinear and very complex, to obtain the joint variables \u03b8j , j = 0, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000642_00368790810902232-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000642_00368790810902232-Figure1-1.png", + "caption": "Figure 1 Schematic of a hydrodynamic lubrication journal bearing with a simple slip wedge", + "texts": [ + " Using the limiting shear stress model and quadratic programming algorithm (Wu and Sun, 2006), we study the load support capacity and friction drag of an infinite length hydrodynamic lubrication journal bearing with a slip wedge. The results show that such an infinite length journal bearing gives an optimization performance differing from either the infinite length slider bearing or the finite length journal bearing. A schematic of a hydrodynamic lubrication journal bearing with a simple slip wedge is shown in Figure 1. In the figure, e is the eccentricity, f is load support angle, v is angular velocity, u is angular coordinate, w is fluid load support, Ra and Rb are the radii of shaft and sleeve, respectively. The shaft surface is supposed to be designed to give a limiting shear stress high enough to inhibit wall slip, but the sleeve surface is an ultrahydrophobic surface exhibiting different slip properties at different regimes. In the regime of 0 , u , uslip, the sleeve surface is designed as ultra- hydrophobic surface with a zero interfacial limiting shear stress (perfect slip surface)", + " The numerical algorithm was described in detail in the literatures (Wu et al., 1992; Wu and Ma, 2005; Wu and Sun, 2006). In our numerical solution, the studied zone was divided by 200 linear elements. We used the same numerical method as given in the literatures (Wu and Ma, 2005; Wu and Sun, 2006) to investigate the hydrodynamic lubrication of the journal Chengwei Wu Volume 60 \u00b7 Number 6 \u00b7 2008 \u00b7 293\u2013298 D ow nl oa de d by A th ab as ca U ni ve rs ity A t 1 3: 12 2 5 Ju ne 2 01 6 (P T ) bearing as shown in Figure 1. The following dimensionless parameters are used: dimensionless pressure P \u00bc pc2/(hvR2), W \u00bc wc2/(hvR3) and dimensionless volume flow Q \u00bc q/(cvR), where c \u00bc Rb 2 Ra ,, R < Rb, q is the fluid volume flow in unit length. In this section, we will study the hydrodynamics of a simple slip wedge as shown in Figure 1. When u , uslip, a perfect slip is allowed on the sleeve surface. No slip is allowed when u . uslip. Figure 2 shows the performance of the journal bearing when the end position of slip zone uslip \u00bc p, p/2 and 0 (no-slip), respectively. At a low eccentricity ratio, the journal bearing with a slip wedge gives a higher load support capacity (Figure 2a), a lower friction coefficient (Figure 2b-c), a smaller load support angle (Figure 2d) and a larger volume flow (Figure 2e) than the traditional journal bearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001544_tmag.2010.2064331-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001544_tmag.2010.2064331-Figure1-1.png", + "caption": "Fig. 1. Problem configuration: bobbin coil in an eccentric tube. For clarity, the coil is shown to be coaxial with the inner surface of the tube. Coils with offset from the main axis are also taken into account.", + "texts": [ + " The potential solution in the different regions of the solution domain is expressed in terms of Fourier series and integrals of the eigenfunctions, with expansion coefficients that are determined by imposing the appropriate boundary conditions across the structure interfaces. The excitation term is treated in a separate paragraph. The part of the theory is concluded with the numerical evaluation of the expansion coefficients. The results of the proposed formulation are then discussed in the following section where verification of the model is done by comparing results to results from FEM. Let us consider an infinitely long tube whose outer cylindrical surface is offset with respect to the inner one by a distance as shown in Fig. 1. The radii of the inner and outer surfaces are and respectivelly, with . The tube is assumed to be conductive and nonmagnetic, and the embedding medium is considered\u2014without loss of generality\u2014to be air. Let denote the conductivity of the tube, whereas stand for the permitivity and the permeability of the medium, which in this case are equal with their respective values in free space (nonmagnetic conductor). The structure is excited by an air-cored cylindrical coil located in the interior of the tube with its axis parallel to the axes of the tube interfaces but it does not necessarily coincide with either of them. Two local cylindrical coordinate systems can be defined, with -axes coinciding with the axes of each tube interface. In the rest of this paper the unprimed coordinates will refer to the coordinate system defined by the inner interface, whereas the primed ones will correspond to the local coordinate system related to the external interface (cf. Fig. 1). The subsequent analysis will be restricted to the time-harmonic regime, i.e., the time convention will be followed throughout this paper. The asymmetry of the structure gives rise to a three-component magnetic field, thus a second order potential ansatz is adopted [12], [17]. According to the latter, the magnetic flux density can be written as (1) where is a second-order vector potential given by (2) and are scalar potential terms which stand for the and solutions, respectively. Both potentials satisfy the Helmholtz equation (3) with ", + " The modal expansion for the potential terms in (7) becomes then (8) (9) where and are the modified Bessel functions. The coefficient is the expansion coefficient for the potential produced by the coil in the absence of the tube, and it therefore depends only upon the coil geometry. Its calculation will be considered in the next section. The potential expressions inside the tube wall ( and ) read (10) (11) where . The solution in the exterior of the tube reads (12) We recall here that the position vectors correspond to the two local coordinates systems depicted in Fig. 1. The potentials in the interior region and the tube wall are expressed in terms of the unprimed coordinate system of the inner interface, whereas the primed coordinates are used for the potential expression in the external region. The above choice for the coordinate system is made in order that at least one of the tube boundaries conforms with the -coordinate surface in each case. The coefficients , , , , , and can be determined by imposing the boundary conditions on the two interfaces of the tube, namely the continuity of the normal magnetic flux component and the continuity of the tangential magnetic field components", + " Substituting the potential expressions (8)\u2013(11) and taking the orthogonality of the exponential functions with imaginary arguments into account, the above continuity relations lead to the following equations: (16) (17) (18) The boundary conditions on the outer interface , read (19) for , (20) for , and (21) for . Since and are expressed in the unprimed coordinate system, they have to be first converted to the primed one. This transition can be done using Graf\u2019s addition theorem for Bessel functions, which for the coordinate systems of Fig. 1 reads [19] (22) (23) Substituting in the expressions (10) and (11) and using again the orthogonality of the exponential functions, the continuity relations (19)\u2013(21) give after some manipulations (24) (25) (26) The expression for the excitation term of an air-cored bobbin coil is derived by the integral representation of the azimuthal component of the magnetic vector potential produced in freespace [12] (27) where (28) and the position vector being defined with respect to the center of the coil (cf", + " For a typical bobbin coil, the coil offset cannot be large and, moreover, eddy-current testing is typically applied to thin-wall tubes, i.e., tubes with a wall thickness much smaller that the average tube radius. Thus, appropriate values for and were found to be small and in all cases they were both set to 5. Note that for a well-defined linear system. It is interesting to note that the number defines the size of the matrix that needs to be inverted. This number is , so in our case we solve 30 30 systems for each in the integral calculation. The configuration studied is shown in Fig. 1. In all cases examined, the coil has inner radius , outer radius , height , and is wound with wire turns. The coil inductance can be calculated from (35) and is found to be 943.6 . The tube is made of 304-type austenitic stainless steel (nonmagnetic) having a nominal conductivity of and a constant inner surface radius of . The tube eccentricity is created by changing the outer surface offset with respect to the inner surface. In addition, the coil is also given an offset. Theoretical and numerical results are compared in Table I for two outer tube surface radii ( and 13 mm), for two outer tube surface eccentricities ( and 2 mm) and for two coil position offsets ( and 1 mm)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000871_09544062jmes949-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000871_09544062jmes949-Figure2-1.png", + "caption": "Fig. 2 Typical HD torsional stiffness curve", + "texts": [ + " The FS is a thin-walled cylinder cup fitted with \u2217Corresponding author: Department of Mechanical Engineer- ing, \u00c9cole de technologie sup\u00e9rieure, Universit\u00e9 du Qu\u00e9bec, 1100 rue Notre-Dame oust, Montreal, Canada H3C 1K3. email: zhaoheng.liu@etsmtl.ca teeth on the outer surface of its open side, whereas the closed side of the cup is a thick wall usually connected to the output shaft of the assembly, and theCS consists of a thick ring with teeth on its inner surface; these teeth engage with those of the FS in a hole pattern and allow the CS to be mounted on the housing of the HD. The teeth of the CS do not fully mesh with those of the FS due to the elliptical geometry of the FS once it is assembled with theWG (Fig. 2). Usually, power is transmitted from the WG to the FS, which rotates in the direction opposite to that of theWG, while the CS is maintained at a fixed position. Developed more than 40 years ago, HDs are still being studied by engineers trying to improve their performance. This is due to the complex principle underlying this device, based primarily on the elastic deformation of the FS, which is limited in displacement by the surrounding more rigid components of the HD (i.e. the CS and theWG). The literature on HDs shows that the relationship between the torque and the torsional angle of the drives due to elastic deformation is non-linear [1\u20135]. Figure 2 shows a typicalHD torsional stiffness curve. JMES949 \u00a9 IMechE 2009 Proc. IMechE Vol. 223 Part C: J. Mechanical Engineering Science by guest on January 20, 2015pic.sagepub.comDownloaded from For applications requiring high positioning accuracy, this non-linear stiffness behaviour should be considered, and control systems may be required to improve their performance [1\u20133]. This behaviour can be traced back to the deformation of the HD components when subjected to significant torques. It has been demonstrated through experiments and the finite-element model (FEM) that the contact areas between the teeth of the FS and the CS increase with the applied torque [6,7]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001946_j.engfracmech.2011.07.012-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001946_j.engfracmech.2011.07.012-Figure4-1.png", + "caption": "Fig. 4. FE geometry and mesh refinements.", + "texts": [ + " Note that the magnitude of the spherical contact parameters depended on the position of the cylindrical contact, the contact geometry in Table 1 and the material properties in Table 2. The finite element (FE) program ABAQUS (6.9) was used to numerically determine the evolvement and distribution between the cylindrical and spherical contact loads during an over-rolling load cycle. The FE model consisted of two three-dimensional bodies: a right cuboid solid with an asperity on its otherwise flat upper contact surface and a section of a cylindrical roller, representing the pinion and follower, respectively, see Fig. 4a. The thickness of the FE model was t = 1/3 mm. Due to symmetry, only half of the geometry was modelled and plane strain was enforced throughout the y-direction, i.e. at both y = 0 and y = t. The dimensions of the model were larger than approximately ten times a characteristic dimension of the contact zones: al in the x and z-directions and ap in the y-direction. Comparison with a three times thicker model showed less than 4% difference in the contact pressure. The mesh in Fig. 4a contained 73,949 eight-node brick elements with several mesh refinements at the contact surfaces and around the asperity. Fig. 4b represents a close-up view around the asperity. Three-dimensional transition elements were used for the 3-refinement [14]. The element size at the centre of the asperity in Fig. 4b was approximately 1.4 lm. The FE simulation consisted of two steps. Firstly, the roller was indented into the flat surface with a prescribed vertical displacement in order to get a maximum contact pressure equal to p0 l. Secondly, the roller was subjected to a prescribed rotation and the cuboic body to a horizontal translation. The rotation and the horizontal translation were related through a constraint equation, which ensured rolling with 3.7% negative slip. Large deformations kinematics were used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure1-1.png", + "caption": "Figure 1: Full toroidal Variator schematic", + "texts": [ + " The result is a family of cost-effective, small, power dense variable drive units that can be applied to ancillary powertrain systems to improve the overall powertrain efficiency. Two applications are described here namely : 1. Taking the \u201ctwo roller\u201d design concept from the OPE application and developing simple variable drive systems for superchargers and 2. Using the Variator technology from the mechanical hybrid as a variable drive unit for a turbo compounding system. *9-2009-01-1465* The heart of the Torotrak variable drive system is the full toroidal traction drive Variator. The schematic in Figure 1 explains the operating principle of the Variator. The power source (normally an engine) drives the input discs (1) and power is transmitted via the rollers (2) to the output discs (3). When the rotational velocities of the input and/or output discs change, the rollers automatically alter their inclination in order to adjust to the new operating conditions (4). Power transmission is achieved by traction, i.e. by shearing an extremely thin, elasto-hydrodynamic fluid film (traction fluid [1]) and not through metal-to-metal friction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003725_icuas.2015.7152366-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003725_icuas.2015.7152366-Figure2-1.png", + "caption": "Fig. 2. Perspective view, body-axis definition and MAVion actuation inputs.", + "texts": [ + " Section II delineates the physical and mathematical model of MAVion and comments 978-1-4799-6009-5/15/$31.00 \u00a92015 European Union 816 over some design issues. Section III describes the process of parameter identification and the wind tunnel campaign that supported this work. Section IV portraits the control strategy adopted and rigorously prove stability for the whole hover-tohorizontal flight transition. Finally, concluding remarks and perspectives are given in section V. MAVion contains 4 moving parts with respect to the aerodynamic fuselage as Fig. 2 illustrates. Roughly, two elevons with deflections \u03b41 and \u03b42 deliver pitching moment (with respect to body-fixed axis y\u0302b) and rolling moment (with respect to body-fixed axis x\u0302b) by means of symmetrical and asymmetrical deflections, respectively, while two propeller engines with rotation speeds \u03c91 and \u03c92 deliver thrust and yawing moment (with respect to body-fixed axis z\u0302b) by means of symmetrical and asymmetrical rotation speeds. Notice that is the case because the motors are installed in a counter-rotative tandem configuration with rotation direction chosen to counter wing tip vortices, which artificially increases the aspect ratio" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001927_s12541-012-0021-7-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001927_s12541-012-0021-7-Figure1-1.png", + "caption": "Fig. 1 A tribometer to measure the friction coefficient between nylon6 and steel", + "texts": [ + " First, a tribometer is set to measure the friction coefficient between nylon6 and steel according to the normal contact pressure. Thereafter, the measured friction coefficient is mathematically converted to efficiency. In order to covert the normal contact pressure into the output torque of a worm gear, gear geometry and Hertz\u2019s Law11 are employed. Through these procedures, worm gear efficiency is predicted. In this study, a worm gear pair which consists of a steel worm and a nylon6 worm wheel was used. In order to measure the friction coefficient between nylon6 and steel, a tribometer was set as shown in Fig. 1. A nylon6 ball which is fully fixed by a screw cap is at the end of a flexible bar. Steel disk which is made of SUM43 is under the nylon6 ball and rotated by a motor. A load cell fixed at the end of a rigid bar is connected to the side wall of the flexible bar so that the load cell can measure the friction force between the steel disk and the nylon6 ball. Weights are used to provide normal load on the disk. In order to obtain the relationship between the friction coefficient and the normal load, experiments were performed with various weights" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure12-1.png", + "caption": "Fig. 12 Schematic of experimental apparatus", + "texts": [ + " For the SWG tested in vacuum, however, the damage at the wave generator\u2013flexspline interface was the most remarkable. In this section, the effect of the environmental pressure on the lubricating conditions of SWG was investigated. The cause of the shortened life in vacuum operations is discussed later. The lubricating condition of SWG was characterized by the lubricant film formation between moving mechanical parts of SWG. For this work, the contact electric resistance technique was applied. The experimental apparatus is shown in Fig. 12 [3]. A test SWG was assembled inside the test jig. The load torque to the output shaft was applied by the rotating arm placed outside the vacuum chamber. Output voltages for monitoring the lubricating conditions were derived outside the chamber by the rotating shafts and two slip rings. The environmental pressure in the vacuum chamber was changed between 3 \u00d7 10\u22123 and 105 Pa by laboratory air purging. The SWG and its lubricant used in the experiment were the same as those in the development test" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001117_045104-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001117_045104-Figure9-1.png", + "caption": "Figure 9. Eccentricity of supporting points of centers to the rotational axis.", + "texts": [ + " The gear checker outputs the difference between the theoretical tooth form and the real tooth form of the measured object as a measurement result but it is affected by these error factors. The uncertainty sources consist of error factors. The VGC simulates the gear measurement considering the following six error factors shown in figure 8: (a) Eccentricity of upper and lower centers. The supporting points of centers are ideally on the rotational axis of the gear checker. However, the actual center has eccentricity to the rotational axis, and it is assumed that the supporting point rotates around the rotational axis as shown in figure 9. This eccentric motion of the supporting point deviates the axis of the measured object from the rotational axis and, thus, the measured object precesses during the measurement. The supporting points are expressed as vectors Ru and Rl in the three-dimensional space. (b) Rotation angular error. The rotation angle of the measured object is measured using a rotary encoder. The rotation angular error \u03b8 rot due to the rotary encoder or the structure of the rotation system affects the measurement result" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002606_s00707-012-0799-5-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002606_s00707-012-0799-5-Figure6-1.png", + "caption": "Fig. 6 Optimal path of the mobile manipulator. a Optimal trajectory of Scout robot (spatial view), b optimal trajectory of Scout robot (upper view)", + "texts": [ + " So, as the robot moves, the output data resulting from two encoders of wheels and two potentiometers of joints are read, and the simulation and experimental results are compared with each other. Figures 5 and 6 show the comparison of the optimal displacements of first and second joints and the optimal paths between the simulation results and the results obtained by experiment for the Scout mobile robot, respectively. The final error of the mobile base is very diminutive and about 0.0053 m, and the final error of the end effector is a bit more and about 0.0153 m. As it can be seen from Fig. 6, the simulation results can be followed by experimental tests, while the moving obstacle avoidance is done in both cases. Also, as it can be seen from Figs. 5 and 6, results obtained by the experiment are effectively followed by those obtained by the simulation study. So, it can be concluded that employing the presented approach the robot can plan the optimal path by considering the distance to the obstacle in every moment, and the mobile robotic system can move among the moving obstacles in the real environment in very efficient way" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000958_iccas.2010.5670241-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000958_iccas.2010.5670241-Figure2-1.png", + "caption": "Fig. 2 Mecanum wheel", + "texts": [ + " There are 3 methods for implementation of omnidirectional driving which are swerve, holonomic and Mecanum drive (see Fig. 1). The swerve drive is consists of independently steered drive modules. It is a simple concept but complex to program and control. The holonomic drive is consists of omniwheels. It is simple to program and control but it has low traction. The Mecanum drive is consists of Mecanum wheels. It is simple to build and program, control. In this paper, we use four Mecanum wheels to omnidirectional driving. The Mecanum wheel was invented by the Swedish engineer Bengt Ilon in 1973. Fig. 2 shows Mecanum wheel. It consists of a set of k congruent rolls placed symmetrically around the wheel body. The face of each roll is part of a surface of revolution R whose axis b is skew to the wheel axis a. Generally, an angle \u03b4 between a and b of \u00b145\u00b0 is chosen. However the Mecanum wheel unlike general wheel is difficult to analysis kinematics so that accurate control is hard. We propose driving control method using fuzzy inference system for mobile robot with Mecanum wheel[1-2]. This paper divided into five sections", + " (5) ~ (8), the following equations are obtained: 1 ( ) w X Y z v v L WV (9) 2 ( ) w X Y z V v v L W (10) 3 ( ) w X Y z V v v L W (11) 4 ( ) w X Y z V v v L W (12) Combining Eqs. (9) ~ (12) into Eq. (13). 1 2 3 4 1 1 ( ) 1 1 ( ) 1 1 ( ) 1 1 ( ) w X w Y w z w V L W v V L W v V L W L WV (13) The radius of a wheel is Rw. Then, the forward kinematics equation is derived. 1 2 3 4 4 w X R v (14) 1 2 3 4 4 w Y R v (15) 1 2 3 4 4( ) w Z R L W (16) By compounding each wheel's force vectors we can move the mobile robot (see Fig.2). For example, right side movement is possible with forward rotating of n.2, n.3 wheel and reverse rotating of n.1, n.4 wheel. Similarly, forward right diagonal side movement is possible with forward rotating of n.1 and n.4 wheel. However, if slip occurred during the movement, the mobile robot moves in an unexpected direction due to unbalanced forces[3-4]. We adjudge the mobile robot to be slipped when a rotation angle from kinematics and an estimated rotation angle do not correspond. Then each wheel's velocity is controlled by the difference of these angles to correct orientation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure12.8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure12.8-1.png", + "caption": "Fig. 12.8 Two types of robot platform currently used on the Teleworkbench (a) Mini-robot Khepera II with some extension modules (from left clock-wise): ASIC with associative memory, a gripper and a linear camera module, Bluetooth module, VGA camera module, and infra-red communication module. (b) BeBot minirobot developed by the Heinz Nixdorf Institute has more computational power than Khepera II", + "texts": [ + " Thus, it is possible, for instance, to use big robots, such as Pionner3-DX, as the robotic platform and a basketball field as the experiment field. The following subsection shortly presents the two types of robot that are currently used in our group. Khepera II [8] is the successor of the older version of miniature mobile robot Khepera. It is designed with functionality similar to larger robots used in research and education. There are some advantages of Khepera II, some of which are: compact, easy to use, decent microprocessor, many sensors and actuator extensions (see Fig. 12.8a showing the Khepera II with some extension modules). Moreover, it is widely used for real-world testing to support developing algorithms, namely trajectory planning, obstacle avoidance, pre-processing of sensory information, and hypotheses on behavior processing BeBot is a new minirobot platform developed by HNI [9]. The robot has a base area of about 9 x 9 cm with a height of about 7 cm (see Fig. 12.8b). It has chain drives with DC motors and incremental encoders which enable it to operate on rough surface. For processing a PXA270 microprocessor at 520 MHz running Linux and a Spartan3E 1200 FPGA for reconfigurable computing are integrated. The robot is equipped with 64 MByte SDRAM and 64 MB Flash RAM. For communication, Bluetooth and Zigbee are integrated by default, while WLAN can be connected via a USB slot. Integrated connectors include USB, RS232, IC, SPI, and MMC/SD Card. The sensors of the robot include 12 infrared-sensors with a range of up to 10 cm and a color camera with a resolution of 640 x 480 pixels" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure10-1.png", + "caption": "Fig. 10 Arrangement of coordinate systems Sf, Sn, and Sp in case of uniform type of taper", + "texts": [ + " Comparison with ystems Sm2 and system Sb in Fig. 3 b allows the following achine-tool settings to be obtained. Em2 = 0 29 XB2 = 0 30 XD2 = 0 31 m2 = R2 32 The blade angles are approximated to the pressure angles at eference point P in expect of further investigation. g1 1 33 g2 2 34 The ratio of gear roll is given by see Eq. 8 m2c2 = 2 c2 = 2 p = cos 2 sin 2 35 4.2 Derivation for Uniform Type of Taper. In case of a piral bevel gear drive with uniform type of taper, the dedendum ngles 1 and 2 are equal to zero. Figure 10 shows new location f systems Sp, Sn, Sf, and S2. The root line is parallel to the pitch ine, wherein is located reference point P. The shortest distance etween the two lines is b2. The derivations made for standard ype of taper are similar for uniform type of taper, making angle 2=0. Sliding base in this case is XB2=\u2212b2. Machine-tool setings are given by S2 = A2 + R2 2AmRu sin 36 ig. 9 Location of the blades and reference point P in coordiate system SP r2 m u ournal of Mechanical Design om: http://mechanicaldesign" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.22-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.22-1.png", + "caption": "Fig. 2.22 The model of the spring detent escapement. a Front view b Back view", + "texts": [ + " Due to the virtual absence of sliding friction between the escape tooth and the pallet during impulse, the spring detent escapement could be made more accurate than lever escapements. Unfortunately, the spring detent escape was rather fragile, not self-starting and harder to manufacture in volume. In 1805, Earnshaw and Arnold\u2019s son (by then John Arnold was deceased) were awarded by the Board of Longitude for their contributions to chronometers. Earnshaw was also known for his bimetallic temperature compensator, and Arnold simplified the complicated structure of the chronometer by applying a helical balance spring. The spring detent escapement is shown in Fig. 2.22. It consists of an escape wheel, a roller with an impulse pallet and a locking pallet, as well as a detent made of a blade, a horn and a spring. The operation of the spring detent escapement is shown in Fig. 2.23. In Fig. 2.23a, before the first shock, the impulse roller rotates counter clockwise and the horn of the detent locks the escape wheel. Figure 2.23b shows the first shock. The impulse roller contacts the detent causing the first shock and pressing it down allowing the escape wheel to move forward" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001497_j.matdes.2010.02.029-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001497_j.matdes.2010.02.029-Figure7-1.png", + "caption": "Fig. 7. Variation in scan spacing and hatch pattern on single layer.", + "texts": [ + " Images of stress patterns for each layer similar to RP process are stored as jpg images using FEM software. These images have been used to compute contour information for each sub part and written in the form of CLI (common layer interface) file which has been used as an input to RP machine for part fabrication. Different subparts have been fabricated with different process parameters to vary mechanical properties in different regions with some overlapping area to ensure strength of the joint between subparts. This enables user to apply different process parameters for single layer (Fig. 7). Developed methodology can reduce weight of parts by applying low energy density and faster prototyping by applying higher laser velocity and scan spacing in non-critical regions obtained by CAE analysis. Further the method can be used in design and fabrication of functionally graded materials. A case study of crane hook is presented here to describe implementation procedure. Typical stress distribution for crane hook is obtained using FEM analysis and presented in Fig. 8. Cross-sectional images of stress levels are retrieved at each section along the Z direction similar to the build direction in RP process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001190_icelmach.2010.5608143-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001190_icelmach.2010.5608143-Figure5-1.png", + "caption": "Fig. 5 Analysis model in this study.", + "texts": [ + " To solve the Maxwell's equation, the magnetic vector potential A is employed, and in this study, the eddy current is ignored. Then the fundamental equation is as follows: M\u0391 rot)rot(rot 0\u03bd\u03bd = (25) where v and v0 are the reluctivity of magnetic material and vacuum, respectively, M is the magnetization. Transmission torque is calculated by the nodal force method. The gear ratio of the analysis model is decided in accordance with (6). In this study, the specifications are shown in Table II, and the dimensions are shown in Table III. The analysis model is shown in Fig. 5. In this model, 14 pole pieces are connected by the flux path facing to the highspeed rotor, whose width is 0.5 mm, equal to the thickness of the laminated silicon steel sheet. This structure helps not only to assemble, but also reduce the cogging torque. However, it is thought that the maximum transmission torque slightly decreases due to the short-circuit magnetic flux. V. TRANSMISSION TORQUE ANALYSIS A. Analysis Condition The maximum transmission torque is generated as the phase difference between two rotors is 90 degrees" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure3-1.png", + "caption": "Fig. 3 Geometry of the channels", + "texts": [ + " Therefore, only 30% of the circumferential surface can be used for the abrasive layer (related to the height of the channels). In the area of the emersions, with only 30% of abrasive layer, high wheel wear can be expected, resulting in low tool life. To increase the tool life of the GIC, the area for the abrasive layer had to be enlarged. Therefore, the geometry of the outlets had to be adapted. It was necessary to create a transition from the channel geometry inside the grinding wheel (inner geometry) to the geometry at the circumferential of the grinding wheel (outer geometry) (see Fig. 3). The coolant outlets are rotated by a setting angle relating to the centerline of the wheel. High setting angles result in a better supply of the contact zone: they increase the areas that are covered by outlets and thus by direct coolant supply (see Fig. 4). However, high setting angles are difficult to machine because of the undercuts in the transition area of the outer and inner channel geometry. As a compromise between good coolant supply and feasible undercuts, a setting angle of 45 degree has been specified" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002339_iros.2011.6095049-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002339_iros.2011.6095049-Figure3-1.png", + "caption": "Fig. 3. Sub-mechanism and kinematic model of RO, with the end effectors set as reference coordinates (x = 0, y = 0 at end effector).", + "texts": [ + " The trajectory of pelvis is provided as A pair of robotic orthosis are connected to PA and parallel attached to lower limbs of subject. The hip, knee and ankle joints of subject can obtain the active assisted, while the metatarsal joints are passive spring connection without control. xm = Xh -l3sin(Bhip) - l4sin(Bknee -Bhip) + lscos(Bknee -Bhip+Bankle) Ym = Yh -l3cos(Bhip) - l4COs(Bknee -Bhip) + lssin(Bknee -Bhip+Bankle) (4) (5) where l3 = 40 em, l4 = 40 em and ls = 15 em, -450 < Bhip < 450, 00 < Bknee < 600, and -300 < Bankle < 300 , the kinematic model is shown in Fig. 3. The overground walking requires intermittent ground contact. During the contact, it is necessary to provide smooth impact between the foot and the ground at initial heel contact. After this, foot-ground contact conditions are needed to keep out the foot dragging within the whole stance phase. Furthermore, requirements of single and double foot stance phase are provided when coordinate both the right and left foot. Finally, foot clearance requirement are provided in the swing phase. These are the issues especially crucial to subjects over a course of rehabilitation process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003695_s00170-015-7533-0-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003695_s00170-015-7533-0-Figure2-1.png", + "caption": "Fig. 2 Simplified RPS diagram with conventional mobile inkjet print head. 1 Solvent reservoir. 2 Solvent feed. 3 Print head. 4 Extending rollers. 5 Ribbon roll. 6 Supporting ribbon. 7 Upper bunker. 8 Excess powder remover. 9 Lower bunker. 10 Solvent reservoir. 11 Solvent feed. 12 Print head. 13 Support powder upper bunker. 14 Excess support powder remover. 15 Support powder lower bunker. 16 Pressure rollers.", + "texts": [ + "ru 1 Institute of Physical Materials Science of the Siberian Branch of the Russian Academy of Science, Ulan-Ude, Russia The solution with one-layer easily dissolved, compressible ribbon by inkjet technology perforation is provided to increase the perforation accuracy from 1200\u00d71200 dpi up to 9600\u00d7 2400 dpi (with several levels of greyscale achieved by controlling drop size (Fig. 1) [18]) and to make the component roll of about 1.5 l in volume (with the ribbon about 5 \u03bcm in thickness after pressure roller work) in about an hour. In case of using about 3-\u03bcm ribbon, it is possible to make an object of about 1 l in volume. Figure 2 shows a simplified view of the inkjet RPS machine with mobile print head. First, a ribbon roll is rewound in the arrow\u2019s direction by extending the rollers (4) and the 17 Component roll Fig. 3 a\u2013e Component ribbon perforation, filling and compression supporting ribbon (6). While the ribbon is being rewound, it is preheated for better perforation in front of the inkjet print heads (3 and 12). Then, drops make tiny holes in the places where a powder (e.g. aluminum with melting point about 660 \u00b0C or PZT ceramics with sintering temperature above 1200 \u00b0C) needs to be poured (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000114_j.wear.2007.01.078-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000114_j.wear.2007.01.078-Figure2-1.png", + "caption": "Fig. 2. Schematic view of the rol", + "texts": [], + "surrounding_texts": [ + "A s l s b n t d s o \u00a9\nK\n1\na o o p o [\ni c a s c l\n0 d\nWear 263 (2007) 663\u2013668\nShort communication\nEffect of specimen preparation on contact fatigue wear resistance of austempered ductile cast iron\nC. Brunetti, M.V. Leite, G. Pintaude \u2217 Postgraduate Program in Mechanical and Materials Engineering, Contact and Surface Laboratory, Federal University of Technology \u2013 Parana\u0301,\nAv. Sete de Setembro, 3165 Curitiba/PR, Zip code: 80230-901, Brazil\nReceived 1 September 2006; received in revised form 10 January 2007; accepted 16 January 2007 Available online 23 May 2007\nbstract\nAustempered ductile iron (ADI) has been tested as a wear resistance material for applications where the applied loads are very high and cyclic, uch as gears. This paper analyzed the endurance lifetime of this material in wear testing equipment that applies contact fatigue stresses in a ubricated ball-on-flat system. The tests were performed at 3.0 or 3.7 GPa of maximum Hertz pressure, using ISO 46 lubricant at 85 \u25e6C, until palling occurrence. The effect of specimen preparation was studied for ground or polished specimens. The worn surfaces were characterized y means of an optical and electron microscope and by the difference between unworn and worn surface profiles. It was found that the graphite odules were exposed at the surface in two ways: cracked and partially exposed. These morphologies were not found in polished specimens, but he wear process in all test conditions produces them. The low endurance lifetime observed in ground specimen was explained based on these efects, which were present before the tests. The slope of Weibull curves was related to the width of the worn track. For polished specimens the lope was smaller than that observed for ground conditions. Thus, any manual process for preparation can be considered as forbidden in order to btain better results.\n2007 Elsevier B.V. All rights reserved.\nurance\np d a s\nl a\ne c\n2\neywords: Contact fatigue; Austempered ductile iron; Surface roughness; End\n. Introduction\nA mechanical system is usually composed of elements that re in contact and under loading. Wear can take place as a result f this contact after a certain period of time. A particular mode f wear is that caused by contact fatigue, which occurs in comonents subjected to cyclic pressures such as gears. This mode f wear is the main cause of failure in these kind of components 1].\nDuring the design of these components the reduction in severty of operational conditions is desired. A way to reduce the ontact fatigue wear is the appropriate materials selection. Usully, ultra cleaned and hardened steels are used in components ubjected to high contact pressures. Nevertheless, nowadays omponents such as gears have been made using steel with high evel of alloying elements, following heat treatment. Another\n\u2217 Corresponding author. Tel.: +55 41 3310 4660; fax: +55 41 3310 4660. E-mail address: pintaude@utfpr.edu.br (G. Pintaude).\nf T n b p c\n043-1648/$ \u2013 see front matter \u00a9 2007 Elsevier B.V. All rights reserved. oi:10.1016/j.wear.2007.01.078\nlifetime\nossibility is the use of cast irons, especially the austempered uctile iron\u2014ADI. Although the graphite nodules could be act s stress raisers, ADI has shown a satisfactory performance when ubjected to contact fatigue wear [2].\nThe surface roughness has a huge impact on endurance ifetime of components subject to contact fatigue [3]. The ustempered ductile iron has been used in some researches [4\u20136].\nThis paper analyzes the effect of specimen preparation on ndurance lifetime of austempered ductile iron, tested in a lubriated ball-on-flat system.\n. Experimental procedure\nAustempered ductile iron (ADI) was subjected to contact atigue wear and its chemical composition is presented in able 1. The bars with 95 mm of diameter and 45 mm of thickess were produced through continuous casting process. These ars were austenitized at 910 \u25e6C for 1.5 h followed by austemering in a salt bath at 290 \u25e6C for 2 h. The metallographic haracterization was performed by following the ASTM A247", + "664 C. Brunetti et al. / Wear 263 (2007) 663\u2013668\nS m\na n 5\nw e t (\ng w a\nc e A T m a i p l i\nb F\nb a i T 0\ntandard [7] procedures. Fig. 1 presents the as-received ADI icrostructure revealed by the optical microscope. The volumetric fraction of graphite nodules is 10.0 \u00b1 0.6% nd their average diameter is 23 \u00b1 10 m. The bulk hardess of ADI is 362 \u00b1 12 HB2.5/187.5 and its bainitic matrix has 08 \u00b1 28 HV0.05.\nThe bars were turned in order to manufacture specimens with asher geometry having the following dimensions: 55 mm of\nxternal diameter, 30 mm of internal diameter and 5.5 mm of hickness. Two initial surface conditions were produced: ground G) and polished (P).\nThe ground surfaces were obtained in two different plane rinding machines, both using an aluminum oxide grinding heel, grade AA-100 G5 VF8 (particle size with 0.16 mm of verage diameter). The processes were differentiated by the\nw a a i\nope: (a) graphite nodules distribution and (b) bainitic matrix.\nontrolling of the longitudinal feed: in one of them, this paramter was not controlled, while in the other it was 7.9 m/min. lso, in both processes the transverse feed was not controlled. he other parameters used in grinding were the same for both achines: an in-feed of 10 m, a total depth of cut of 50 m and speed of 30 m/s. The average roughness of ground specimens s around 0.04 m. After grinding process, the polishing was erformed manually with 1 m diamond grains, using metalography technique. The average roughness of polished surface s about 0.01 m.\nThe rolling contact fatigue (RCF) tests were performed in all-on-flat equipment, projected and constructed by Leite [8]. ig. 2 presents a schematic view of the used equipment.\n1.2 l of lubricant oil ISO 46 are circulated through the chamer where specimens are located. The tests were always started fter the oil temperature reached 85 \u25e6C, and this temperature s maintained by electric resistance during all period of tests. he employed velocity of 1700 rpm allowed a load frequency of .5 \u00d7 105 cycles/h. The counter-body was an AISI 52100 sphere\nith 5/16 in. of diameter and it was received with 0.01 m of\nverage roughness. The contact pressure \u2013 3.0 or 3.7 GPa \u2013 is chieved by means of dead-weights applied under a cantilever; n order that an axial load is employing on three balls placed\nling contact fatigue test rig.", + "ear 263 (2007) 663\u2013668 665\ni fl w t t c\nc 3 w\nc m t c\ne d\na o s n s c p c u c r\ne m r w F w T m\n3\nc i o i P t\nF f\nt c\nd c m s t s o s c a\nw s\nm c n s l F t a g w\nT R\nT\nP G P\nC. Brunetti et al. / W\nn a cage of a 52206 thrust bearing. The design of the ball-onat system allows a load application limited to 150 kg of mass, hich results in 1.75 kN of maximum axial force, considering he action of the cantilever. In order to obtain 3.0 GPa of conact pressure 59 kg of mass were applied, while for 3.7 GPa test ondition, 119 kg of mass were disposed as dead-weights.\nEach test condition was identified by the manufacturing proess (grinding or polishing) and by the contact pressure (3.0 or .7 GPa). For example, P3.7 means that the test was conducted ith polished specimens under a 3.7 GPa of contact pressure. The number of cycles to the failure was associated to sound hanges, perceived by the operator. After all tests a spalling echanism was detected. The fatigue life was evaluated through he Weibull distribution of two parameters [9]. Each testing ondition corresponded to five specimens of ADI.\nThe worn surfaces were analyzed using optical and scanning lectron microscopes. The depth and width of worn tracks were etermined through surface roughness measurements.\nThe removal of graphite nodules was determined by metllographic counting before and after RCF tests, using areas f 0.05 mm2, outside and inside the worn tracks. For polished pecimens, in each counting, it was identified that the empty odules were caused by the removal of the graphite. For ground pecimens, due to the smaller amount of removed nodules, the ounting was performed considering the kinds of morphologies roduced by the grinding process, classified in this paper as racked subsurface nodules (CSN) and partial subsurface nodles (PSN). All average values of graphite nodules counting orresponded to one hundred measurements for each studied egion.\nThe surface roughness was determined in SURTRONIC 25+ quipment. The evaluation length was 4 mm. The stylus diaond was perpendicularly disposed to the worn surface. The esulting profiles were analyzed in the TALY PROFILE softare \u22123.1.10 version, to calculate the Rpk roughness parameter. or each roughness profile, the depth and width of worn track as determined, in four positions, each one separated by 90\u25e6. he average values for each position corresponded to fifteen easurements.\n. Results and discussion\nFig. 3 presents the failure probability as a function of the load ycles. From Fig. 3 it is possible to extract the results presented n Table 2, which summarizes the L10, L50 and average values f contact fatigue life. Table 2 shows that the endurance lifetime ncreases as the contact pressure was reduced, i.e., \u03b7 value of 3.7 condition is smaller than those observed for P3.0 or G3.0 ests. More than that, the lifetime of ground specimens is smaller\nb t i w\nable 2 olling contact fatigue tests results\nest condition L10 (cycles \u00d7 106) L50 (cycles \u00d7 106)\n3.7 0.05 0.13 3.0 1.02 1.48 3.0 1.70 2.78\nig. 3. Failure probability as a function of load cycles for ADI tested in the ollowing conditions: ( ) P3.7, ( ) G3.0 and ( ) P3.0.\nhan the observed value for polished ones, tested at the same ontact pressure (3.0 GPa).\nThe reduction shown in Table 2 for endurance lifetime ue to the initial surface roughness was 1.89. Dommarco and o-workers [4,10] observed reductions in the same order of agnitude \u22121.79 \u2013 when they compared ground and polished pecimens of ADI in RCF tests. The observed difference for he endurance lifetime is due to the modification on the contact tress distribution. The rough contact causes local disturbances n the Hertzian pressure distribution, producing stationary presure spikes [3,11]. This effect makes easy the nucleation of racks, which is responsible by up to 85% of the total life of component subjected to the contact fatigue [11].\nThe morphology of graphite nodules exposed at the surface as very different, when one compared the ground to polished pecimens. Fig. 4 presents these aspects before RCF tests. Fig. 4 shows that in the ground specimen there were two\norphologies of graphite nodules. These morphologies were lassified as cracked subsurface nodules and partial subsurface odules. All nodules in the Fig. 4a are positioned in the suburface, because the grinding process gives rise to a metallic ayer over them. On the other hand, some points presented in ig. 4b could be classified as PSN, and it was not observed he CSN morphology in polished specimens. Probably, the metllographic preparation eliminated the CSN morphology. For round specimen, the number of occurrences of CSN and PSN as determined and the results are presented in Table 3. Table 3 shows that the amount of PSN morphology is smaller efore RCF tests than that observed inside the worn track, since he CSN/PSN ratio decreases after wear process. This result is an ndicative that the deformed and cracked layer over the nodules as removed during tests and a large amount of nodules became\n\u03b7 Medium life \u03b2 Weibull slope r2\n0.16 2.03 0.94 1.60 4.94 0.93 3.03 3.91 0.93" + ] + }, + { + "image_filename": "designv11_3_0002117_10402004.2011.629404-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002117_10402004.2011.629404-Figure7-1.png", + "caption": "Fig. 7\u2014Motion of contact relative to the disc.", + "texts": [ + " Even in the middle of the stroke, where the track has been previously depleted of lubricant, it can be seen that menisci formed at the side of the contact due to the oil pulled in from the side bands. In the presence of a lateral movement there is an additional mechanism of replenishment that consists of a thick layer of grease in the side bands being pushed back onto the track by the changeable direction of motion of the ball relative to the disc. Due to the lateral oscillatory motion, the actual path of the ball on the disc surface deviates from the initial worked track, as seen in Fig. 7. Depending on the ratio between the rotational frequency of the disc and the frequency of the lateral motion, the contact may not pass through exactly the same point on the disc for many revolutions of the latter. It has also been found that, at least under the conditions in these tests, the recovery of the film thickness under lateral oscillations was permanent and lasted for the duration of the tests, which is longer than the 5 min shown in the graph in Fig. 6. D ow nl oa de d by [ N ew Y or k U ni ve rs ity ] at 1 5: 23 0 7 D ec em be r 20 14 Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002761_icppw.2012.69-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002761_icppw.2012.69-Figure1-1.png", + "caption": "Figure 1. (a) The yaw angle \u03c8; (b) The pitch angle \u03b8; (c) The roll angle \u03d5.", + "texts": [ + " Euler angles include the yaw angle, denoted by \u03c8, pitch angle, denoted by \u03b8, and roll angle, denoted by \u03d5. The yaw angle (respectively, pitch angle and roll angle) is the magnitude of the rotation by the z-axis (respectively, y-axis and x-axis) in the counterclockwise manner, i.e., in the direction from the x-axis to the y-axis (respectively, from the zaxis to the x-axis and from the y-axis to the z-axis). Let B1 = {x1,y1, z1} and B2 = {x2,y2, z2} respectively denote the standard bases of the pre-rotation and postrotation coordinate systems. Fig. 1(a), 1(b) and 1(c) respectively illustrate the rotations by a yaw angle, a pitch angle and a roll angle. In the case of the rotation by a yaw angle, illustrated in Fig. 1(a), we have x2 = cos\u03c8x1 + sin\u03c8y1 = [ cos\u03c8 sin\u03c8 0 ]T B1 , y2 = \u2212 sin\u03c8x1 + cos\u03c8y1 = [ \u2212 sin\u03c8 cos\u03c8 0 ]T B1 , z2 = z1 = [ 0 0 1 ]T B1 . The coordinate transformation matrix TB2\u2192B1 , conventionally denoted by Y\u03c8, is Y\u03c8 = cos\u03c8 \u2212 sin\u03c8 0 sin\u03c8 cos\u03c8 0 0 0 1 . Therefore, we have [v]B1 = Y\u03c8 [v]B2 for every vector v. Similarly, in the case of the pitch rotation, y f y f q illustrated in Fig. 1(b), and the roll rotation, illustrated in Fig. 1(c), the coordinate transformation matrix TB2\u2192B1 of the pitch rotation and the roll rotation, conventionally denoted by P\u03c8 and R\u03d5 respectively, are P\u03b8 = cos \u03b8 0 sin \u03b8 0 1 0 \u2212 sin \u03b8 0 cos \u03b8 and R\u03d5 = 1 0 0 0 cos\u03d5 \u2212 sin\u03d5 0 sin\u03d5 cos\u03d5 . The orientation change of an object can be achieved by a sequence of rotations of Euler angles. The YawPitch-Roll convention (\u03c8, \u03b8, \u03d5) in which a yaw rotation is followed by a pitch rotation and then a roll rotation is used in this study, and the corresponding coordinate transformation matrix is Y\u03c8P\u03b8R\u03d5 = c\u03c8c\u03b8 \u2212s\u03c8c\u03d5+ c\u03c8s\u03b8s\u03d5 s\u03c8s\u03d5+ c\u03c8s\u03b8c\u03d5 s\u03c8c\u03b8 c\u03c8c\u03d5+ s\u03c8s\u03b8s\u03d5 \u2212c\u03c8s\u03d5+ s\u03c8s\u03b8c\u03d5 \u2212s\u03b8 c\u03b8s\u03d5 c\u03b8c\u03d5 where c is shorthand for cos and s is shorthand for sin" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003472_j.cirpj.2011.01.010-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003472_j.cirpj.2011.01.010-Figure10-1.png", + "caption": "Fig. 10. Ring-on-disc test.", + "texts": [ + " Hence, the hydrodynamic condition could likely be reached more rapidly by use of this patterned structure. Patterned spherical-segment microstructures with the above combination of geometric parameters were fabricated and evaluated in a tribometer as described in the following sections. In order to measure the hydrodynamic effect of microstructures, a sufficiently large surface area is necessary. For this reason, it was decided to build a tribometric ring-on-disc test rig for the experimental investigations (Fig. 10). Only the ring counterpart featured the surface microstructures. The discs for the tribological tests consisted of heat treatable steel 42CrMo4 (1.7225) and had an outer diameter of dsa = 100 mm and a thickness of hs = 8 mm. The ring counter bodies were produced of bronze (2.1030) with an outer diameter dra = 100 mm, inner diameter dri = 70 mm and thickness hr = 10 mm. Further details are listed in Table 2. Patterned spherical segments were produced by Jet-ECM. In this process, a high-velocity electrolytic jet is generated from a cathodically polarized nozzle and led toward the anodically polarized workpiece" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001975_s12555-010-0301-x-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001975_s12555-010-0301-x-Figure3-1.png", + "caption": "Fig. 3. Ball and beam systems (internal synchronization).", + "texts": [ + " 3) Controlled synchronization: it is the result of external actions (control inputs), such as master-slave system. Controlled synchronization can be classified as two types: [2] 1) Internal synchronization: all systems are synchroniz- ed in the same level. 2) External synchronization: a system is more important than the other in multiple systems and his movement can be regarded as independent of the movement of other systems, for example, master-slave system. In this section we discuss the synchronization of two types of ball and beam systems which is shown in Fig. 3. It is internal synchronization. For the \u201cBall and beam 2\u201d, the gravity of the beam can be neglected. We define synchronization error s as 1 2 1 2 ,s r r s r r= \u2212 = \u2212 r1 and r2 are the positions of the ball in \u201cBall and beam 1\u201d and \u201cBall and beam 2\u201d, see Fig. 3. The synchronization control problem is to design a controller which computes the applied voltages U1 and U2 such that the synchronization error s reaches zero. In this paper, we only discuss serial PD synchronization control. For the parallel PD synchronization control, similar results can be obtained. We design a serial PD synchronization control as ( ) ( ) ,i pmi i i dmi i i pbi dbi iU k k k s k s\u03b1 \u03b1 \u03b1 \u03b1 \u03b2 \u03b2 \u03c0 \u2217 \u2217 = \u2212 + \u2212 + + + ( ) ( ), 1,2,d d i pbi i dbi ik r r k r r i\u03b1 \u2217 = \u2212 \u2212 \u2212 \u2212 = where i \u03c0 is synchronization compensator, 0\u03b2 > is a synchronization constant, pmik and dmi k are positive gains for the motor control, pbik and dbi k are positive gains for the ball control, i \u03c0 is the compensator for each system, d r is reference position for the two balls" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000691_s0025654408030059-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000691_s0025654408030059-Figure2-1.png", + "caption": "Fig. 2.", + "texts": [ + " On this plane, the stationary motions are associated with the set \u03a3 = \u03a3+ \u222a \u03a3\u2212 \u222a \u03a30, where the \u03a3\u00b1 are the straight lines q = \u22131 + p2 2(1 \u2213 b)2 (q = f(\u00b11)), (16) the set \u03a30 consists of curves determined parametrically by the formulas (see (14), (15)) q = f(x, p2), p2 = \u03d5(x), (17) and x \u2208 (\u22121, 1) is a parameter that should be eliminated from relations (17). The shape of the Smale diagrams significantly depends on the top parameters a and b. On the basis of a detailed analysis of the function \u03d5(x), the plane of these parameters is divided into seven domains (a)\u2013(g) (Fig. 1) each of which is associated with its own Smale diagram (Fig. 2 a\u2013g). Domains (a)\u2013(g) are determined by the relations (a) a > 1 + b; (b) 1 + b > a > \u03b1+(b); (c) \u03b1+(b) > a > \u03b1\u2217(b); (d) \u03b1\u2217(b) > a > \u23a7 \u23a8 \u23a9 1 \u2212 b, b \u2265 7 \u2212 \u221a 48, \u03b1+(\u2212b), b < 7 \u2212 \u221a 48; MECHANICS OF SOLIDS Vol. 43 No. 3 2008 (e) \u03b1+(\u2212b) > a > 1 \u2212 b; (f) 1 \u2212 b > a > \u03b1\u2212(\u2212b); (g) a < \u23a7 \u23a8 \u23a9 1 \u2212 b, b \u2265 7 \u2212 \u221a 48, \u03b1\u2212(\u2212b), b < 7 \u2212 \u221a 48. Here \u03b1\u00b1(b) = 1 \u2212 b 8 (7 + b \u00b1 \u221a 1 + 14b + b2), \u03b1\u2217(b) = (1 + 3b2)(1 \u2212 b2) 1 + 6b2 + b4 . On the Smale diagrams (Fig. 2 a\u2013g), the bold lines distinguish the curves \u03a3\u00b1 and \u03a30 or their parts corresponding to local minima of the effective potential, i.e., to stable steady motions. The unmarked curves \u03a3\u00b1 and \u03a30 or their parts correspond to local maxima of the effective potential, i.e., to unstable steady motions. The bifurcation values of the dimensionless constant of the Jellett integral, at which stability is changed and the steady motions are branching, are determined by the relations p2 \u00b1 = (1 \u2213 b)4 a \u2213 (1 \u2212 b) , p2 \u2217 = 16a3/2(1 \u2212 a \u2212 b)3/2 3 \u221a 3(1 \u2212 b)2 . (18) In particular, if the top parameters satisfy the condition 1 \u2212 b < a < 1 + b [1, 2], then (Fig. 2 b\u2013e) the top fast rotations about the vertical symmetry axis are unstable for the lowest position of the center of mass (p2 > p2 +, i.e., \u03c92 > \u03c92 +) and stable for its highest position (p2 > p2 \u2212, \u03c92 > \u03c92 \u2212). Here (see (13) MECHANICS OF SOLIDS Vol. 43 No. 3 2008 and (18) we have \u03c92 \u00b1 = mgc C (1 \u2213 b)2 a \u2213 (1 \u2212 b) . We note that all points belonging to the set \u03a3 on the plane (p2, q) are invariant with respect to the MECHANICS OF SOLIDS Vol. 43 No. 3 2008 phase flow of system (1)\u2013(4) for M = 0, and all other points evolve along the line p2 = const in the direction of decreasing q. This fact allows us to perform global qualitative analysis of the dynamics of a dynamically symmetric ball with displaced center of mass on a plane with sliding friction. For example, let the top parameters belong to the domain (b) (Fig. 1); in this case, the generalized Smale diagram has the form shown in Fig. 2 b. If we set the top in a position close to the stable equality \u03b3 = e and let it spin fast (p2 = p2 0 > p2 \u2212 > p2 +) about its symmetry axis, then on the plane (p2, q) such initial conditions are associated with a point lying in a neighborhood of the line \u03a3+ to the right of the line p2 = p2 \u2212. If this point lies below (above) the line \u03a3+, then its starts to move along the line p2 = p2 0 in the direction of decreasing q and uniquely (with probability 1) tends to the point p2 = p2 0 lying on the line \u03a3\u2212; i", + " Thus, for almost all initial values, the final motions of the top whose parameters lie in domain (b) (Fig. 1) are uniform rotations about the vertical symmetry axis for the lowest (if p2 0 < p2 +) or the highest (if p2 0 > p2 \u2212) position of the center of mass or precession (if p2 0 \u2208 (p2 +, p2 \u2212)) motions. But if the top parameters lie in the domains (c)\u2013(e) (Fig. 1), then, just as in the above case, the fast rotation of the top about the vertical symmetry axis is stable (unstable) for the highest (lowest) position of the center of mass (fig. 2 c\u2013e). But, in contrast to the case considered above, one can predict the final motions of the top whose parameters lie in domains (c)\u2013(e) with probability 1 only for p2 0 < p2 \u2217 or p2 0 > p2 + (in cases (c) and (d)) and for p2 0 < p2 \u2212 of p2 0 > p2 + (in case (e)). If these inequalities are violated, it is impossible to predict the final motions of the top whose parameters lie in domains (c)\u2013(e) with probability 1, because for the same value p2 0 \u2208 (p2 \u2217, p 2 +) (in cases (c) and (d)) and p2 0 \u2208 (p2 \u2212, p2 +) (in case (e)), there exist two stable steady motions of the top", + " Obviously, the friction torque M destroys the steady motions (12), but under conditions (20), system (1)\u2013(4) admits quasisteady motions close to steady ones. Therefore, the above qualitative analysis of the top dynamics without the friction torque taken into account can be considered as the generating case of analysis. Consider the top whose parameters lie in domain (b) (Fig. 1). We set the top in stable equilibrium or in a state close to it and let it spin fast about the symmetry axis (p2 0 > p2 \u2212 + \u03b4, \u03b4 > 0). Such initial conditions are associated with point Q on the plane (p2, q) (Fig. 2 b) lying in a neighborhood of the line \u03a3+ on the vertical p2 = p2 0. Under the action of the phase flow of the system (1)\u2013(4), this point starts to move in the direction of decreasing q (\u201cfast\u201d) and p2 (\u201cslow\u201d) until it comes to the line \u03a3\u2212 to the right of the line p2 = p2 \u2212 (if \u03b4 > 0 is sufficiently large). Then point Q starts to move \u201cslowly\u201d along the line \u03a3\u2212 downwards and to the left. Coming (for p2 = p2 \u2212) to the point of bifurcation at which the line \u03a3\u2212 is tangent to the curve \u03a30, point Q starts to move \u201cslowly\u201d along the curve \u03a30 to the left and downwards" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000135_s1560354708040096-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000135_s1560354708040096-Figure1-1.png", + "caption": "Figure 1", + "texts": [ + " 4 2008 Assume that the friction force linearly depends on the normal force: F = \u03b1(q, q\u0307) + \u03b2(q, q\u0307)N (2.4) where \u03b1, \u03b2 are vectors lying in the tangent plane. Substituting (2.4) in (2.3) and taking the complementarity condition into account, we obtain \u039b\u0308 = { U, if N = 0, U\u2217 + V \u2217N, if N > 0, U\u2217 = U + (W,\u03b1), V \u2217 = V + (W,\u03b2) (2.5) Conclusions on the contact or detachment can be made by considering the signs of the coefficients U , U\u2217, and V \u2217 in formula (2.5). The results of this analysis are presented in Fig. 1 and in the table. Table. Types of motion in the system with a unilateral constraint depending on the coefficients of Eqs. (2.3) and (2.5): \u2191 means detachment, \u2194 means contact, \u2191\u2194 means that both these types are possible, and \u2205 means that both these types are impossible. U\u2217, V \u2217 1 2 3 4 ++ \u2212+ +\u2212 \u2212\u2212 U A + \u2191 \u2191\u2194 \u2191\u2194 \u2191 B \u2212 \u2205 \u2194 \u2194 \u2205 On the plane of the variables N and \u039b\u0308, by the complementarity condition, only points lying on the positive coordinate semiaxes are admissible (the so-called corner law). The graph of function (2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003304_1.3616914-FigureI-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003304_1.3616914-FigureI-1.png", + "caption": "Fig. I Schematic v i e w of system being a n a l y z e d", + "texts": [], + "surrounding_texts": [ + "A. E S H E L Department of Mechanical Engineering, Columbia University and Ampex Corporation, Redwood City, Calif. Assoc. Mem. ASME\nH. G. E L R O D , J R . Lubrication Research Laboratory,\nDepartment of Mechanical Engineering, Columbia University, New York, N. Y.\nAssoc. Mem. ASME\nStiffness Effects n Hi Infinitely f i le Foil Bearing Numerical solutions are presented for the film thickness of the infinitely wide, selfacting foil bearing for various values of tape stiffness. The solutions indicate that with increasing stiffness there are: (a) a slow reduction in the nearly uniform clearance prevailing under most of the wrap angle; (b) an increase in the peak of the undulations occurring in the region where the tape leaves the spindle; and (c) virtually no change in the trough of these undulations.\nI introduction I HE behavior of a thin foil when transported over a\nspindle is of interest in many fields Of particular interest is its relation to a magnetic tape traveling over a read-write head. The situation under consideration is depicted in Fig. 1. In a recent publication [l],1 solutions for the foil contour were obtained under the assumptions of one-dimensional flow, fluid incompressibility, and negligible foil stiffness. It is the purpose of this paper to show how the foil stiffness may be incorporated in the formulation of the problem and to determine its effect quantitatively.\nBasic Equations The foil is considered as an infinitely wide plate loaded by a pressure p(6) and by tensions applied far from the spindle. Consequently, the plate assumes the shape of a cylindrical surface with variable radius of curvature R(0). Since the foil thickness I is very small compared with the radius of curvature, and since the mode of deformation is that of an infinitely wide developable surface, the usual small-deformation theory of plates is applicable.\nThe clearance is defined by the equation\nh = r -\nd / dp\\ dh Te r i - le\n(2a) 2\n-Nomenc lature-\ndP rr (/l ~ h\"> (2b)\nForced and moment balances on a small element of tape require;\ndT Q b \u2014 + T = 0 ds R\n(1) V _aU2 T _ dQ\nVa~ ~ R + R ds where r is the radial polar coordinate, and ro is the radius of the spindle.\nThe Reynolds equation of lubrication with errors of order (h/n)\ndM ds\n0 = 0\n(4 )3\n(5)\nThe symbol T represents the fluid shear traction on the foil (positive if acting in the +s-direction)\nh dp U\nor equivalently 1 Numbers in brackets designate References at end of paper. 2 When the symbols \u00b1 , T appear, the upper sign applies to the case where the independent variable is taken as positive in the direction of motion, and the lower sign where it is not.\nContributed by the Lubrication Division and presented at the Lubrication Symposium, New Orleans, La., June 5-9, 1966, of THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS. Manuscript received at ASME Headquarters, March 10, 1966. Paper No. 66\u2014 LubS-6.\nT\n2)' 0 de \"\nt \" M h (6)\nand the moment M per unit width is related to the radius of curvature by\nJ f - 4 R\n(7)\n3 The result of the centrifugal force is effectively a reduction in tension by aW1. Hereafter, the symbols T, To,, and so on, will stand for this effective tension.\na = integration constant h = clearance tote to head A = integration constant K = distance of asymptote to head L = see Fig. 1 b = integration constant h* = nominal clearance (at loca- m, n = exponential measures B = integration constant tion dp/dO = 0) M = bending moment per unit c = integration constant H = dimensionless clearance = width C = integration constant (h/r\u00bb)6 M = normalized moment D = flexural rigidity of tape per H* = dimensionless nominal clear- V = pressure under foil unit width ance = (/i*/r0)e~2/ '3 Va = ambient pressure E = modulus of elasticity H = normalizedclearance = H/H* 1 = real root of characteristic } = small perturbation = h/h* equation F = integration constant H = R - 1 Q = = shear force per unit width G = integration constant Ha = normalized distance of asymp- (Continued on next page)\n9 2 / J A N U A R Y 1 9 6 7 Transac t ions of the ASiVIE Copyright \u00a9 1967 by ASME\nDownloaded From: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jotre9/28537/ on 03/10/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "where D is the flexural rigidity per unit width [4]:\nD - Et\u00b0 ~ 12(1 - v2)\nThe element of length ds along the foil is related to the polar angle by\n1 ds 1 \u201e .. _ IS = ~To {r'dd' + dri) n dO r- 0, one finds\nm = i n = D/Tafo2 ~ 0 (e ! / ' ) (19)\no o\nHence a stiffness parameter of order 1 may be defined:\nS = De~*/>/Tar02 (20) It is well to take a look at the physical meaning of the foregoing operations. D/ro represents, according to (7), the moment required to bend the foil to a radius of curvature r0. Hence D/T^r02 is the ratio of this moment to the moment of the tension about the center of the spindle. Three cases may be distinguished: (a) When the foregoing ratio is of smaller order than e2^', say of order e 0 is the drag coefficient. The functions cL and cD are called aerodynamic characteristics of the body.\nThe control design is much simplified when Fa does not depend on the vehicle\u2019s attitude \u03b8. For example, a constant velocity flight is achieved by aligning the thrust direction R(\u03b8)e1 in (1) with the external force (mge1 + Fa), via the angular velocity control \u03c9, and by setting T = |mge1 + Fa| (modulo correction terms needed to compensate for tracking errors w.r.t. a given reference trajectory). Enforcing this strategy when Fa depends on \u03b8 is far from obvious because any change of the vehicle\u2019s thrust direction (i.e. the vehicle\u2019s orientation) affects the external force as well. However, we show in the next section that, for a specific class of aerodynamic characteristics, this latter case can essentially be recast into the former case.\nProposition 1 Assume that the resultant of the aerodynamic forces is of the form (5).\n(i) If the aerodynamic characteristics are given by\n{ cD(\u03b1) = c1 + 2c2 sin2(\u03b1),\ncL(\u03b1) = c2 sin(2\u03b1), (6)\nwith c1 and c2 denoting two constant real-numbers, then the change of thrust control input\nT \u2212\u2192 Tp = T + 2c2ka|x\u0307a| 2 cos(\u03b1 \u2212 \u00b5), (7)\ntransforms the system\u2019s equation (1) into:\nmx\u0308 = \u2212TpR(\u03b8)e1 + mge1 + Fp(x\u0307a, \u00b5), (8)\nwith\nFp(x\u0307a, \u00b5) = ka|x\u0307a| [ cL(\u00b5)S \u2212 cD(\u00b5)I ] x\u0307a (9a)\ncD(\u00b5) = c1 + 2c2 cos2(\u00b5), (9b)\ncL(\u00b5) = \u2212c2 sin(2\u00b5). (9c)\n(ii) If the aerodynamic characteristics cD and cL are re-\nspectively even and odd functions (as in the case of symmetric bodies), i.e. {\ncD(\u03b1) = cD(\u2212\u03b1), cL(\u03b1) = \u2212 cL(\u2212\u03b1), (10)\nthen (6) is the only family of aerodynamic characteristics for which there exists a function Fp independent of \u03b8 such that (8) holds true for any \u00b5 \u2208 S 1.\nThis proposition points out the modeling functions (6) for the aerodynamic characteristics used in the remainder of this paper and the fact that, in the case of symmetric bodies, they are the only functions which allow for the transformation of the system\u2019s equation (1) into (8) with Fp independent of the attitude \u03b8.\nGiven (3), (4), (5), (6), (7) and (9), it is possible to prove the item i) of the proposition by verifying via direct calculations that \u2212TRe1+Fa(x\u0307a, \u03b8, \u00b5) = \u2212TpRe1+Fp(x\u0307a, \u00b5). As a matter of fact, Proposition 1 is a corollary of a more general result stated below. It addresses the problem of transforming System (1) into the form (8), when Fa is given by (5) without the symmetry constraints (10) being imposed upon the aerodynamic characteristics.\nTheorem 1 Assume that the resultant of the aerodynamic forces is given by (5). Then,\n(i) System (1) can be transformed into the form (8) with\nFp independent of \u03b8 if and only if the aerodynamic characteristics cL(\u03b1) and cD(\u03b1) satisfy the following differential equation:\n(c\u2032\u2032D \u2212 2c\u2032L) sin(\u03b1 + \u00b5)+(c\u2032\u2032L + 2c\u2032D) cos(\u03b1 + \u00b5) = 0. (11)\n(ii) If (11) holds true, the vector valued function Fp de-\npending only on x\u0307a and \u00b5, and the scalar function Tp such that (8) holds true are respectively given by:\nFp(x\u0307a, \u00b5) = ka|x\u0307a| [ cL(\u00b5)S \u2212 cD(\u00b5)I ] x\u0307a (12)\nwith {\ncD(\u00b5) = cD(0) + c\u2032L(0) cos 2(\u00b5)+c\u2032D(0) sin(\u00b5) cos(\u00b5), cL(\u00b5) = cL(0) \u2212 c\u2032D(0) sin 2(\u00b5)\u2212c\u2032L(0) sin(\u00b5) cos(\u00b5),\nand\nTp = T + ka|x\u0307a|x\u0307 T a R(\u03b8) [ \u2212c\u2032L(\u03b1) c\u2032D(\u03b1) ] . (13)\nThe proof of this theorem is given in the paper\u2019s appendix.\nNote that condition (11) is satisfied for any value of \u00b5 only if {\nc\u2032\u2032D \u2212 2c\u2032L = 0 \u2200\u03b1, c\u2032\u2032L + 2c\u2032D = 0 \u2200\u03b1. (14)" + ] + }, + { + "image_filename": "designv11_3_0002182_gt2010-22440-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002182_gt2010-22440-Figure1-1.png", + "caption": "Fig. 1 Photograph of prototype metal mesh foil bearing", + "texts": [ + " The paper presents measurements of (externally induced) forced response in a MMFB afloat a flexible overhang rotor and details a simple identification procedure for estimation of bearing force coefficients, frequency dependent. Impact loads are employed along only one direction; and assumes for small amplitude motions about a centered position that the principal force coefficients are identical (KXX=KYY) and the cross coupled coefficients are asymmetric (KXY=-KYX). The simplification follows from the lightly loaded condition of the test bearing. The MMFB, as depicted in Figure 1, employs commercially available ring shaped metal mesh as the elastic support under a top arcuate foil. Table 1 details the dimensions of the bearing. The metal mesh ring is a compressed weave of copper wires (0.30 mm in diameter). The mesh is 20% in compactness, offering large structural energy dissipation characteristics [19]. The top foil, a smooth arcuate surface 127\u00b5m thick, wraps the journal when not in operation. The top foil is made from a cold rolled steel strip (Chrome-Nickel alloy, Rockwell 40/45) with significant resilience to deformation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002528_c2sm25387a-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002528_c2sm25387a-Figure1-1.png", + "caption": "Fig. 1 (a) diagram of actuation testing apparatus; (b) schematic illustration of the procedure for carrying out constant strain actuation measurements at increasing strains. Each \u2018\u2018run\u2019\u2019 consisted of an initial pre-strain adjustment at pH 3 followed by a small amplitude strain oscillation to determine modulus after which the pHwas changed to pH 8 and the small amplitude strain oscillation repeated.", + "texts": [ + "16 After crosslinking, the fibres were pinned in place on the bars of the TEM grid with UV-curable glue which was applied with a glass microfibre, positioned with a micromanipulator, (Marzhauser Wetzlar DC-3 K) while being observed under an optical microscope. Along the length of the fibre, some portions were left uncovered by glue to allow for imaging with the AFM to determine the fibre cross-sectional area. Actuation performance was determined by measuring the change in tensile force in stretched nanofibres when the pH of the surrounding solution was changed. The fibres were stretched by lateral movement of the fully-immersed AFM cantilever, using the method similar to that described previously20,21 and illustrated in Fig. 1a. The fibre was initially immersed in pH 3 solution and stretched at its midpoint to a strain that was sufficient for the fibre to remain under tension once it was eventually swollen at pH 8. The pH was then cycled from pH 3 to pH 8, followed by a pause, then back again to pH 3. Changes in pH were affected by infusing 30 ml of the appropriate solution through the petri dish (total volume of 10 ml) with a push/pull syringe pump operating at 3.5 ml min 1. At each pH the fibre was also subjected to a small amplitude strain cycle to determine the elastic modulus", + " The cantilever was scanned back and forth over 500 nm (equivalent to a strain of approximately 1%) at a constant This journal is \u00aa The Royal Society of Chemistry 2012 500 nm s 1 for a minimum of 5 cycles. Once the pH had been cycled and the 500 nm scans carried out, the fibre strain was then incrementally increased by 2.5%. The pH cycling and 500 nm cantilever scans were then repeated in the same manner at each strain increment. For ease of expression, this procedure (pH cycling and 500 nm scans) carried out at a single strain will be referred to as a \u2018run\u2019 and is schematically illustrated in Fig. 1b. The lateral twist of an AFM cantilever was used as the force sensor for the actuation measurements. The calibration of the cantilever force to lateral deflection ratio (N/Vl), or \u2018lateral force conversion factor\u2019 (al), was determined by the use of a glass fibre as a reference beam.28 The fibres are assumed to undergo purely axial tensile stretching rather than bending during the three point mechanical test used here as is known to occur with low Young\u2019s modulus ( 10 GPa) fibres of high aspect ratio of length to width (>30 : 1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001514_tmag.2009.2018676-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001514_tmag.2009.2018676-Figure2-1.png", + "caption": "Fig. 2. Specification of structure and operation.", + "texts": [ + " However, the performance of spherical wheel motor is not analyzed easily with two different coordinates systems. Therefore the rotor coordinates system should be transformed to the stator coordinates Manuscript received October 16, 2008. Current version published May 20, 2009. Corresponding author: J. Lee (e-mail: julee@hanyang.ac.kr). Digital Object Identifier 10.1109/TMAG.2009.2018676 system as using Z-Y-Z Euler transformation matrix [2]\u2013[4]. Operation of the spherical wheel motor is understood as using directions of alpha degree and beta degree like Fig. 2. The magnets are arranged with regular alpha 0 degree and beta 90 degrees on the rotor. The coils are separated as two layers positioned at alpha 18 degrees, because the position of coil layer is important factor to decide stability of holding torque. The coils in each layer are arranged with regular beta 60 degrees. 0018-9464/$25.00 \u00a9 2009 IEEE Motion of spherical wheel motor is decided by current function which is defined as reference position alpha degree and beta degree. Therefore, the current function is like (1) (1) where ; (upper layer), 1(lower layer); vertical direction angle on spherical coordinates system; horizontal direction angle on spherical coordinates system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002647_0954406212461326-Figure13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002647_0954406212461326-Figure13-1.png", + "caption": "Figure 13. the structure and FEM model of cracked gear.", + "texts": [ + " The finite element mesh model is divided in accordance with space eight-node (hexahedron) element, then, boundary conditions and external loads are applied to analyze the dynamic characteristics of the cracked gear structure, as well as to simulate the influence of different sizes and different locations of cracks on dynamic characteristics of gear structure. The influence of crack can be investigated by comparing to the dynamic characteristics of gear without crack. The results of modeling and mesh dividing are shown in Figure 13. The input parameters are: inner-circumference radius a\u00bc 0.1m, tooth number Z\u00bc 80, gear modulus m\u00bc 10mm, diameter of pitch circle of gear d\u00bc 0.8m, thickness of gear h\u00bc 0.025m, the Young\u2019 modulus of gear elastic body E\u00bc 2.10 1011N/m2, the material density \u00bc 7860 kg/m3, the Poisson\u2019s ratio \u00bc 0.30, and the meshing force of gear P0\u00bc 100N. The two cracks are set to two teeth which are next to each other, and the positions of Figure 11. First fourth-orders vibration shape of gear tooth (xc\u00bc 0.25 L, \u00bc 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000275_020-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000275_020-Figure2-1.png", + "caption": "Figure 2. Cross section of a Goss (110) [001]-oriented sheet, cut normal to the direction of magnetization, illustrating the principal domain wall types treated. Planar domain walls spanning the sheet take up the minimum energy orientations at angles + $ o skew to the sheet normal [I101 direction. Hybrid walls consisting of two or more planar sections alternately skew may also span the sheet; special cases are the bra and ket walls illustrated. Walls which do not span the sheet may bound wedge domains. Such walls are normally located in (100) planes, but the general case is treated in the text.", + "texts": [ + ") Recently, measurements have begun to be made on ideally oriented single-crystal specimens whose domain structure is observed to conform closely to the slab model (Helmiss 1969, Swift er a1 1974). Agreement with the PB losses cannot be regarded as entirely satisfactory. Refinement of the model is now required to match the improved experimental definition. While the errors incurred by PB in neglecting the finite wall thickness and finite specimen width can be shown to be unimportant (Bishop 1976) the consequences of domain wall bowing and the tilt of the domain wall planes away from the sheet normal (figure 2), as in (110) [OOI] Goss oriented silicon iron (Graham and Neurath 1957 GN hereafter) both need to be taken into account. Domain wall bowing (caused by the magnetic field inhomogeneity resulting from eddy currents) has already been the subject of some theoretical analysis (Bishop 1971, 1973) the results of which compare favourably with Helmiss\u2019 (1969) observations on a (100) [OOl] \u2018picture frame\u2019 single crystal of 3+% Si-Fe. The purpose of the present paper is to quantify the effect of domain wall tilting in the absence of bowing, i", + " The concentration of Skew, braket and wedge walls 293 the eddy currents should therefore be reduced and consequently so should the losses as they depend on the mean-square eddy-current density. Domain wall tilting in Goss oriented material with positive cubic anisotropy has been analysed and observed by GN. The domain wall energy per unit area, y, depends upon wall orientation according to the formula y=yoP(a)=2(AK~)112P(a)=yo [l+---- In (: ~ :$:)] N ~ O (1 t-0.38 sin312 (28)). (1) A represciits the exchange constant, K1 the first anisotropy constant, a r 1 -cos28 sin2 8 and 8 is the angle between the wall plane (assumed to contain the [00l] z axis) and a (100) plane: see figure 2. Thus in (110) [001] material, the wall energy per unit area y is minimized when 8=0, i.e. the wall plane is (100) and tilted 4= 245\" to the sheet normal [I 101. The wall area, however, is of course a minimum when 8 = & 45\" and 4 = 0. The total wall energy, for a wall that spans the specimen thickness, is a minimum when y(0) sec4 is a minimum. For ideally oriented Goss material this occurs at compromise angles t $0, where tan $0 = 0.632874, 40 N 32.33\". This analysis neglects KZ and higher anisotropy coefficients and magnetostrictive effects. Both may be shown to have little effect on wall tilting (GN, Lilley 1950). greatly complicates the situation by permitting the existence of a variety of wall types with almost equal energies. Not only may planar skew walls at both orientations 5 40 exist, but also hybrid combinations of them, consisting of two (or more) plane sections alternately oriented at -t 40 separated by a fairly sharp crease or kink through 240 (figure 2). The energy of a kink is very small in comparison with that of the whole wall, the ratio being at most of the order wall thicknesslsheet thickness. Such hybrid skew walls would be statically metastable only in ideal Goss oriented material, for as the angle $ between the sheet normal and nearest (110) plalie is varied either side of the ideal 45\", the two energy minima for planar skew walls become unequal and indeed at #=35\" only one remains. Thus, unless $=45\" exactly, a kinked hybrid wall is statically unstable against migration of the crease in the direction favouring the growth of the section oriented at the lower effective energy y(0) sec (+), Hybrid walls may, however, be favoured in dynamic conditions", + " The greater eddy current drag on the wall elements deep within the strip should encourage migration of a crease of appropriate sign (once nucleated) from the surface towards the wall centre-a quasi-bowing phenomenon. As the energy difference between the minima is small (ON) such dynamic formation of kinked walls is expected to occur well below the threshold velocity for significant true-wall bowing. Such hybrid kinked walls, consisting of two planar regions meeting at a central crease, will be termed 'braket' walls after the Dirac 'bra' and 'ket' symbols < } that they closely resemble. Another type of kinked wall bounds a wedge domain (figure 2). As remarked by GN the boundaries of the wedge should lie in the minimum y (100) and (010) planes for the balance between orientation and area energies no longer applies. The existence of such a wedge must depend on other energy contributions, particularly magnetostatic. The growth of wedge domains through the sheet thickness provides a well known mechanism (see for example Passon 1963) for the creation of reverse domains bounded by planar walls crossing the specimen thickness, i.e. slab domains", + " Averaging (numerically) over a quarter cycle from J=O to J=Jm and dividing by the 'classical' loss P, (equation 11) yields the mean anomaly factor for cyclic magnetization by the array of parallel skew walls as a function of their spacing ratio, 2L/D, skewness angle $, and the degree of saturation b = Jm/Js. Anomaly ratios calculated in this way for + = 40 = 32.33 O are recorded in table 1 under the 'skew' headings for comparison with the corresponding PB (4 =0) values and results for other models to be discussed below. A wedge domain is bounded by parts of two oppositely iiiclined skew domain walls extending only part way (from y=O to y=a< 0) into the sheet (figure 2). Unless the material is of ideal Goss orientation the two walls will be tilted at somewhat different angles +=tan-l T and +I=tan-l TI, As the wedge expands, its cross section will be assumed to sweep through a set of similar triangles, i.e. and $1 will not change as a increases. In general also the two sides will move with somewhat different (x component) speeds V , VI. The mean eddy current field on the first side H ~ = X w o + H w l where HWO is the field arising from its own motion at speed V and HWI arises from the motion of the other side with speed VI" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002001_1.4007884-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002001_1.4007884-Figure2-1.png", + "caption": "Fig. 2 Thermally induced rotor bend and overhung mass", + "texts": [ + " The journal temperature distribution is calculated by solving the transient heat conduction equation @2T @x2 \u00fe @ 2T @y2 \u00bc qc k @T @t (6) The model is restricted to cases with a very low axial temperature gradient in the shaft, which is assumed to be equal to zero. Equation (6) is solved utilizing four node isoparametric elements for discretizing the problem domain (shaft\u2019s circular cross section). Temperature and flux continuity boundary conditions are assigned to the interface between the journal and lubricant (see Fig. 1). kL @TL @x y\u00bch \u00bc ks @TJ @x y\u00bch (7) TLjy\u00bch\u00bc Tsjy\u00bch (8) 2.4 Flexible Rotor Model With Thermal Bow Induced Imbalance. The asymmetric temperature distribution in the journal causes the shaft to bow as shown in Fig. 2. The bow angle at the journal\u2019s midplane is obtained from [17] b\u00f0t\u00de \u00bc bx \u00fe iby \u00bc a I \u00f0r 0 \u00f02p 0 \u00f0L=2 L=2 T\u00f0r;/; z; t\u00der2ei/dzd/dr (9) where a and I are the thermal expansion coefficient and the second moment of area of the journal, respectively. The flexible rotor is modeled with Timoshenko beam elements including shear deformation. The equation of motion for the flexible rotor bearing system can be written as 011701-2 / Vol. 135, JANUARY 2013 Transactions of the ASME Downloaded From: http://tribology" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003024_128_2011_231-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003024_128_2011_231-Figure7-1.png", + "caption": "Fig. 7 (a) The average local orientation of molecules inside the layer fragment is described by the nematic director n \u2013 average orientation of long molecular axes and the polar director p \u2013 average direction of the molecular tips. y is the angle between the layer normal n, and the nematic director, a is the angle between the tilt plane and the polar director, and the azimuthal angle \u2019 determines the position of the nematic director on the tilt cone, defined by the angle y. Orientation of the smectic layer (departure from the planar layer alignment) is denoted by the angle D. The general orientation of molecules in the layer is obtained by combination of the leaning \u2013 rotation around axis perpendicular to polarization direction (b) and tilting \u2013 rotation around polarization direction (c)", + "texts": [ + " It should be stressed that the crystallographic unit cell angle is not directly related to the inclination of molecular long axis inside the layer fragments. There are known examples of the B1rev phase with an orthogonal crystallographic unit cell in which molecules are tilted inside the blocks [12] as well as the oblique crystallographic unit cell made of layers in which molecules are not tilted [8]. The average local orientation of the molecules can be described by the average orientation of the long molecular axis (this we call the nematic director) and by the average orientation of the molecular tips (the polar director) \u2013 see Fig. 7. The angle y by which the nematic director is tilted with respect to the local smectic layer normal Fig. 5 The typical XRD pattern for the structure with the orthogonal primitive unit cell is characterized by the perpendicular mutual positions of the (10) and (01) signals (the tilt angle) can be obtained by optical measurements, as for smectic phases, by observing the texture of a sample placed between crossed polarizers. For samples exhibiting a fan texture with circular domains, the tilt can be deduced by determining the angle by which the extinction brushes are inclined from the polarizing directions", + " For most bent-core smectics the polar vector is perpendicular to the tilt plane, defined by the layer normal and averaged long axis direction, just as polarization in the ferroelectric rod-like liquid crystalline systems. However, since in the bent-core liquid crystals the polar order is decoupled from the tilt order, the polar director can in general have any direction in space; thus it can also have a non-zero component along the layer normal. This can be achieved by a combination of tilting (rotation around the polar director) and leaning (rotation around the direction perpendicular to the polar director) of molecules (Fig. 7) in the smectic layer. If in the consecutive layers leaning of molecules alternates, i.e., banana tips are slightly above and below the layer midplane, then the density across the layers will have a double layer periodicity visible in XRD as a weak subharmonic of the main \u201clamellar\u201d signal (Fig. 10). Additionally, due to the competition between the density modulation associated with the monolayer and the bilayer periodicity resulting from the general tilt structure, the mass centers (banana tips) can change the positions periodically along the layer, being below or above the geometrical center of the layer (Fig", + " This coupling is responsible for the chiral symmetry breakdown in phases where bent-core molecules are tilted with respect to the smectic layer normal [32, 36]. The second term in (7) stabilizes a finite polarization splay. The third term with positive parameter K \u00f01\u00de np describes the preferred orientation of the molecular tips in the direction perpendicular to the tilt plane (the plane defined by the nematic director and the smectic layer normal). However, if K \u00f01\u00de np is negative, this term prefers the molecular tips to lie in the tilt plane. The last term in (7) stabilizes some general orientation (a) of the polar director (see Fig. 7) which leads to a general tilt (SmCG) structure. The stability analysis [32] of the lamellar (one-dimensional) structure shows that it is stable only for weak coupling between the splay of polarization and molecular tilt. When ~Kp is larger than some critical value, layers start to undulate. Let us first consider the case where the preferred orientation of the polar director is perpendicular to the tilt plane (K \u00f01\u00de np > 0). The spatial variation of the layer normal and the nematic and polar directors is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002887_1.4024547-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002887_1.4024547-Figure3-1.png", + "caption": "Fig. 3 Three-pad FB with radial injection holes at the leading edge groove region", + "texts": [ + " The main purpose of this paper is to evaluate the cooling effectiveness of radial injection cooling of a FB-rotor pair using multiple air injection points along both axial and circumferential directions. The current work investigates radial injection cooling applied to three-pad FBs, and results are compared with axial cooling methods. Unlike the single-pad FB investigated by Radil and Batcho [27], the three-pad FBs can implement radial air injection holes at multiple locations along the circumferential direction, as shown in Fig. 3. The effectiveness of the radial injection was compared with that of axial cooling using the same flow rates. Test Rig Configuration Figure 4 presents the test module configurations that were used to compare radial injection cooling with axial cooling. The test foil bearing is housed inside a split housing, and covers at both ends form plenum chambers on both sides of the test bearing. For axial cooling, cooling air is introduced to one plenum and discharged to the other side. In the radial injection cooling, air is introduced at the center of the housing and forced into a total of nine nozzles formed along the circumferential direction, as shown in Fig. 3. Figure 5 Journal of Tribology OCTOBER 2013, Vol. 135 / 041703-3 Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use shows a photo of the entire test rig, which is similar to the rig used in the characterization of transient thermal behavior of a three-pad FB [22]. The test rig consists of a new 10-kW induction motor, two angular contact ball bearings supporting the test journal shaft, a torque rod with a preloaded piezoelectric load cell, and a loading mechanism for the FB" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000558_jf902816e-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000558_jf902816e-Figure4-1.png", + "caption": "Figure 4. Interaction of different monoacid TAGs with reaction temperature on incorporation (mol %) of oleic acid. The reaction conditions are as follows: substrate mole ratio, 1:1; reaction time, 6 h; and enzyme dosage, 10%.", + "texts": [ + " To determine the effect of reaction temperature on FA selectivity of Lipozyme RM IM, temperatures of 40, 50, and 60 C were tested. It was reported that higher temperatures favor higher yields for endothermic reactions due to the shift of thermodynamic equilibrium toward products. An elevated temperature can also make the operation easy, since a higher temperature increases the solubility of the substrate and decreases the viscosity of the solutions (25). Three-dimensional plots for the interaction of different monoacid TAGs with reaction temperature on incorporation (mol%) of oleic acid are shown inFigure 4. Usage of three different reaction temperatures resulted in significant differences in oleic acid incorporation (P < 0.05) when the binarymixture of oleic acid andT12was used. The incorporation rate change was higher when the reaction temperature was increased from 40 to 50 C compared to from 50 to 60 C. Oleic acid incorporation into all TAGs was increased with the increasing temperatures. Itwas reported that higher temperatures reduce the viscosity of the lipid and consequently increase the substrate and product transfer on the surface or inside the enzyme particles (7)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002612_j.csefa.2013.05.002-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002612_j.csefa.2013.05.002-Figure5-1.png", + "caption": "Fig. 5. Schematic illustration of load distribution: (a) evenly distributed load without any misalignment; (b) uneven distribution of load as a result of misalignment; (c) as a consequence of uneven load distribution, stress level at the roller tips increases causing severe damage near the tip of the roller.", + "texts": [ + " The extent of uneven load distribution on the rollers will depend on the degree of misalignment. This uneven load distribution raises the contact stress towards the edges of the rollers abruptly. This abnormal stress level at the tip of the roller can lead to boundary lubrication situation and direct metal to metal contact between the roller and the cone surface generating wear debris. The damage initiates at the tip of the roller, as the stress concentration is higher at the roller tip due to its size being smaller than the cone diameter. The above proposed mechanism is illustrated in Fig. 5. This is apparent in the damaged roller samples where one end of the roller alone shows severe damage with the other end having only deep scratches. This phenomenon was observed in all the rollers. The scratches may be due to the abrading action of the wear debris (worn parts of the roller and cone surface). The wear debris analysis had confirmed the presence of metallic wear particles along with non-metallic particles. The presence of non-metallic particles indicates dirt/dust entrapment during service" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000166_ls.65-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000166_ls.65-Figure1-1.png", + "caption": "Figure 1. Confi guration of a misaligned compliant journal bearing.", + "texts": [ + " The non-linear fl uid\u2013structure interaction problem is solved numerically by means of an iterative procedure. On the other hand, the fi rst perturbation method and the Lund\u2019s stability criterion16 at the threshold of instability are applied to determine the fl uid fi lm stiffness, and damping coeffi cients and stability limits under small disturbances. We used here a model consisting of a simplifi ed horizontal shaft of mass (2 M) assumed rigid and supported symmetrically by two identical compliant (elastic liner) bearings. Figure 1 presents the geometrical and kinematical details for a misaligned single-layered journal bearing. The journal bearing consists of a rigid shaft rotating at angular velocity \u00f9 in a housing covered with a thin elastic layer (thickness th) of elastic characteristics (E,\u03bd). The shaft and the bearing are separated by a thin incompressible viscous non-Newtonian fi lm. Copyright \u00a9 2008 John Wiley & Sons, Ltd. Lubrication Science 2008; 20: 241\u2013268 DOI: 10.1002/ls Modifi ed Reynolds Equation For a journal bearing lubricated with couple stress fl uid operating in dynamic and isothermal conditions, with a laminar fl ow assumption the modifi ed Reynolds equation is given by:12 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202x G h p x z G h p z R h x h t , , ( ) + ( ) = +6 120 0\u00b5 \u03c9 \u00b5 (1) where G h h h h ,( ) = \u2212 + 3 2 312 24 2 tanh In this equation, is the characteristic length of additives; \u00b50 is the shear viscosity (absolute viscosity); \u03c9 and R are the angular velocity and the journal radius, respectively; p is the pressure; and h is the fi lm thickness", + " p p p p ei t= + +( )0 1 0 1\u03b5 \u03b5 \u03c6\u03b5 \u03c6 \u03b3 (4) h h ei t= + +( )0 1 0 1\u03b5 \u03b8 \u03b5 \u03c6 \u03b8 \u03b3cos sin (5) where p p p p \u03b5 \u03c6\u03b5 \u03b5 \u03c6 = =\u2202 \u2202 \u2202 \u2202 and 0 Substituting equations (4) and (5) into equation (2), and collecting the zero and the fi rst-order terms for \u03b51 and \u03b50\u03c61, we get the following set of linear partial differential equations in p\u03030, p\u0303\u03b5, p\u0303\u03c6: \u2022 zero-order equation \u211c( ) = p h 0 06 \u2202 \u2202\u03b8 (6) \u2022 fi rst-order equations \u211c( ) + + p h p R L z h p z \u03b5 \u03b8 \u03b8 \u03b83 30 2 0 2 0 2 0\u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 cos cos\u03b8 \u03b8 \u03b3 \u03b8 = \u2212 +6 12sin cosi (7) \u211c( ) + + p h p R L z h p z \u03c6 \u03b8 \u03b8 \u03b83 30 2 0 2 0 2 0\u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 sin sin\u03b8 \u03b8 \u03b3 \u03b8 = +6 12cos sini (8) where \u211c( ) = ( ) ( ) + ( ) ( ) \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 \u03b8 \u03b8 G h R L z G h0 0 2 0 0, , \u2202 z Copyright \u00a9 2008 John Wiley & Sons, Ltd. Lubrication Science 2008; 20: 241\u2013268 DOI: 10.1002/ls and G h h h h 0 0 0 3 0 2 3 012 24 2 ,( ) = \u2212 + tanh p\u03030 is the static pressure distribution corresponding to the static equilibrium position of the journal centre (\u03b50, \u03c60), h\u03030 is the distribution of dimensionless static fi lm thickness and p\u0303\u03b5 and p\u0303\u03c6 are the complex pressure derivatives due to the journal motion. Considering shaft misalignment (Figure 1), the dimensionless lubricant fi lm thickness at any position can be calculated using the following formula:17 h z0 0 0 01= + + \u2212( )\u03b5 \u03b8 \u03b4 \u03b8 \u03b2cos cos (9) \u03b40 = d C where \u03b40 is the static magnitude ratio, d is the magnitude of the projection of the entire journal axis (C1C2) onto the mid-plane and \u03b20 is the static angle of misalignment. The variation fi eld for these parameters is: 0 1 0 180 180 0 0 < < \u2264 < \u2212 \u2264 \u2264 \u03b5 \u03b4 \u03b4 \u03b2 m \u00b0 \u00b00 \u03b4m is the maximum possible value of \u03b40 at which contact between the journal and the housing takes place: \u03b4 \u03b5 \u03b2 \u03b5 \u03b2m = \u2212 \u2212( )2 1 0 2 2 0 0 0sin cos (10) Now, \u03b40 can be written in terms of \u03b4m and a parameter Dm as follows: \u03b4 \u03b40 = m mD (11) where Dm is the degree of misalignment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001497_j.matdes.2010.02.029-Figure21-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001497_j.matdes.2010.02.029-Figure21-1.png", + "caption": "Fig. 21. Scanning results of fabricated crane hook part (a) part no. 1 (b) part no. 2.", + "texts": [ + " To validate the dimensions and form of the fabricated parts through developed methodology for free form surfaces, crane hook has been fabricated with three set of parameters as given in Table 1 for three regions, viz 1, 2 and 3 of the part as shown in Fig. 11. For comparison a crane hook is also fabricated with standard parameters for better accuracy as suggested by the machine manufacturer as a single part. Fig. 19 shows the fabricated parts as per details given in Table 1. Fabricated parts were reverse engineered using 3D white light scanner (Fig. 20) and then the obtained point cloud data have been superimposed on the CAD model. The deviation plots of the obtained point cloud data with respect to CAD model have been presented in Fig. 21a and b as coloured/shaded maps for part nos. 1 and 2 as given in Table 1 respectively. Side bar shown in this figure represents the deviations of the geometry from the CAD model. It is clearly evident from this figure that dimensions of fabricated part are close to its solid model. The percentage deviation of the geometry from the CAD model is around \u00b10.6 which is within the acceptable limits of accuracy of SLS parts [17] and also same for both the parts presented in Fig. 21. Table 1 also shows the weight of crane hooks fabricated using different process parameters. While fabricating these crane hooks after dividing in three subparts, the issue of strength of joint requires exploration. In this work it has been found that time lag between laser scanning for three parts on a layer is very less and does not affect the strength at joint. It is similar to sorted scanning strategy (Fig. 2), where a contour is scanned in two steps to save the scanning time. Similar procedure is used in up-down skin strategy (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001557_med.2009.5164716-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001557_med.2009.5164716-Figure1-1.png", + "caption": "Fig. 1. Quadrotor in flight, with an outline of the airflow around it. is the incidence angle and cos and sin .", + "texts": [ + " MATHEMATICAL MODEL The design of the proposed hybrid controller requires an elaborate mathematical model. This model needs to describe quadrotor behavior in climb, descent and forward flight. At the same time the model has to be simple enough to provide fast simulation based experiments with the proposed hybrid controller. The momentum theory of a rotor, also known as classical actuator disk theory, combines three important aerodynamic quantities: \u2013 Rotor thrust, \u2013 Induced velocity (airspeed produced in rotor), , \u2013 Airspeed induced from aircraft movement. Fig. 1 shows an example of a quadrotor moving through the air in horizontal direction. For the airflow in Fig. 1 we can write cos and sin while \u03b2 is the incidence angle. The momentum theory combines these quantities into the following equation: 2 \u00b7 . (1) Basic momentum theory equation (1) has two possible solutions. Both solutions define operation states where the defined rotor slipstream exists. Two solutions are the helicopter branch and the windmill branch and they refer to rotorcraft climb and descent respectively. Unfortunately classical momentum theory implies no steady state transition between the helicopter and the windmill branches", + " A cross section of a blade elemental \u03948 at a distance 8 from the rotor center is given in Fig. 2. This figure shows the blade element in quadrotor climb mode when there is no lateral movement of the quadrotor. For better clarity the angles are drawn larger than they actually are. Quadrotor\u2019s climb and induced speeds, and respectively, tend to reduce the mechanical angle of attack 9: 8 to the effective angle of attack 9;< 8 . We also have to consider linear blade twisting 9: 8 \u03980 ( > ?\u0398@A . Where \u03980 is the mechanical angle at the root of a blade, and \u0398@A is the linear twist angle. Fig 1 and Fig 3 show the lateral airstream which is produced when rotorcraft moves in horizontal XY plane. Generally speaking, as can be seen in Fig 1, the stream consists of a vertical flow B and a horizontal flow which form the total flow C $ B . We continue with an observation of a small rotor blade element \u0394r. It produces elemental lift and drag force, EF E> and EE> respectively. The total rotor lift is derived by summing all the elemental lifts of all the blade elements. Because of the blade rotation, forces produced by blade elements tend to change both in size and direction. This is why we need to find an average elemental thrust of all the blade elements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001114_0951192x.2010.528033-Figure22-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001114_0951192x.2010.528033-Figure22-1.png", + "caption": "Figure 22. PUMA robot serving first milling machine.", + "texts": [ + "5 D ow nl oa de d by [ FU B er lin ] at 0 5: 20 1 4 M ay 2 01 5 It is required that the robot links perform a specified sequence of motions to fulfil a certain sequence of jobs predefined by the user. The robot job can be defined as a simple handling process as it is requested to translate a work piece from a specific location to another. To demonstrate the UNIGRAPHICS NX (UG) simulation capabilities a sequence of robot movements is suggested as follows: . The robot picks up a work piece from the stocking place. . Delivers the work piece to the centre lathe (Figure 21). . Translates the work piece to the first milling machine (Figure 22), after the required cutting time by the centre lathe. . If the first milling machine is already loaded, the robot will translate the work piece from the centre lathe to the second milling machine (Figure 23). . After operations completion, the robot will translate the work piece to the storing place. The main served locations in the FMS for which the robot is requested to reach are defined by their Cartesian coordinates [x, y, z] in the robot base frame R0. Table 5 shows the output values of the Puma robot joints [q] for each of the served locations\u2019 coordinates on the three CNC machines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001384_bf02478609-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001384_bf02478609-Figure3-1.png", + "caption": "Fig. 3 Specimen mounting and extensometer arrangement", + "texts": [ + " KERSHAW (1971 ) investigated a number of methods of gripping tendon specimens, and compared measurements of actual specimen extension with measurements of grip separation. Nearly all the methods of gripping which were investigated resulted in some slip, although most methods enabled the specimens to be broken. The method of gripping the specimen, which reduced movement within the grips to a minimum, consisted of coating the ends of the specimen with Eastman's 910 adhesive and clamping them to plastic end plates with small metal damps (Fig. 3). Although some tests have been conducted on skin in vivo (MILLINGTON et al., 1971 ; KENEDI et al., 1965), most experiments are carried out on excised tissue and a suitable test environment must be provided. Steps should obviously be taken to ensure that the excised tissue remains reasonably fresh, and precautions comparable with those used for transplant tissues should be adopted wherever possible. The behaviour of connective tissue is sensitive to water content. Striking changes in mechanical properties occur if the tissues are allowed to dry excessively, but the difference between the behaviour of moist and fully immersed tissues is not great (KERSHAW, 1971), and the simplest method of providing a controlled environment is to immerse the tissue in normal saline or another suitable isotonic fluid of the correct pH. We have tended to mount each specimen in its own small bath (Fig. 3), so that it can be stored and various forms of chemical treatment carried out without removing the specimen. However, if preferred, standard attachments are available which enable tests to be conducted in any large container (ABRAHAMS, 1967). The simple arrangement shown in Fig. 3 makes no allowance for temperature control, and tests with that arrangement have usually been conducted at room temperature. Temperature control is readily achieved in a large tank, or the specimen can be mounted in a standard tissue bath having a separate outer chamber through which liquid is pumped from a temperature-controlled tank. Extension measurements The simplest and most convenient method of measuring specimen extension would undoubtedly be to measure the separation of the grips, but this is only accurate if there is negligible relative movement between the specimen and the grips", + " Methods which do not require physical contact between the specimen and the instrument are to be preferred, and for the less extensible tissues overall amplifications of about 1000:1 are required. As extensometers which satisfied all these requirements were not readily available, an extensometer (Instron type G51-12) was selected which used electrical strain gauges to sense the deflection of a metal cantilever strip. Used in conjunction with an appropriate amplifier and recorder, this extensometer gave adequate sensitivity and continuous read-out but had to be physically attached to the specimen. Several modifications (Fig. 3) were made to the extensometer and are mentioned here because they indicate some of the problems which arise from the need for direct attachment. Stainless-steel extension legs were fitted to enable the extensometer to be kept clear of the Medical and Biological Engineering July 1974 513 fluid in the bath; specially shaped lightweight knife-edge clips were used to reduce local loads near the knife edges; a light closing spring was fitted to prevent the cantilever spring arms from stretching the specimen when the extensometer was attached; the extensometer was counterbalanced to render it effectively weightless, and an attachment was fitted which counteracted the forces required to open the extensometer and reduced the overall stiffness of the instrument to 8 '8 N/m, which was approximately 5 % of the stiffness of the smallest and most extensible specimen (bovine ligamentum nuchae) at its mean strain (40%)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003348_tie.2012.2205353-Figure18-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003348_tie.2012.2205353-Figure18-1.png", + "caption": "Fig. 18. Stator winding with turn-to-turn faults.", + "texts": [ + " In a future publication, the derived model will be used to accurately predict the fault current from the observed fault signature. Using a database generated by this model, the severity of the fault can be evaluated at any operating condition of the generator. A wellinformed postfault protection strategy will then be proposed to prevent the stator winding insulation from further thermal degradation while preserving as much as limp-home capability of the claw-pole generator as possible. APPENDIX The derivation of the equivalent circuit in Fig. 3 is shown in this appendix. Fig. 18 shows a physical representation of the generator winding with the turn-to-turn fault. The lower coil consisted of m turns of the phase-A winding shorted through a fault resistor Rf . The upper coil contains the remaining N \u2212m turns. Since \u03d5A is defined as the magnetic flux that links the entire phase-A winding, it will induce back EMFs in both the shorted coil and the nonshorted coil, as shown by the upper two voltage sources in Fig. 18. As for the shorted coil, assume that the magnetic flux produced by the fault current if either enters the rotor as part of the air-gap flux or only links the shorted coil itself. With this simplified assumption, the effect of the leakage flux produced by the fault current if can be represented as an inductive voltage drop in the shorted coil, as shown at the bottom of Fig. 18. The phase voltage vab is the sum voltage of the two coils vab = (N \u2212m) d\u03c6A dt \u2212 N \u2212m N RsiA +m d\u03c6A dt \u2212m N Rs(iA + if )\u2212 m2 N2 Lls dif dt . (A1) An equation similar to (A1) was derived for permanentmagnet machines. Readers may refer to [12, Eq. (7)] for details. Combining and reorganizing terms in (A1) yields N d\u03c6A dt \u2212RsiA = vab + m N Rsif + m2 N2 Lls dif dt . (A2) In the short-circuit path, the fault current if can be calculated as m d\u03c6A dt \u2212 m N Rs(iA + if )\u2212 m2 N2 Lls dif dt = Rf if . (A3) Reorganizing terms in (A3) yields m N ( N d\u03c6A dt \u2212RsiA ) \u2212m N Rsif\u2212 m2 N2 Lls dif dt =Rf if " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003500_1.3591525-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003500_1.3591525-Figure3-1.png", + "caption": "Fig . 3 T y p e d i a g r a m", + "texts": [ + " To determine the possibility of unlimited rotation of the input crank, we simply interchange r with g, and / with g in the foregoing procedure, obtaining gcr and a range of d-values (d,,d2, d3,di) for a given value of q. Hence, if both cranks are to have unlimited rotations, b must lie between the two intermediate d-values corresponding to the particular values of q and r, both of which must exceed their critical value. Type Determination A survey of types can now be obtained as follows. For a given orientation of the crank axes and offsets (f,y), we can com- pute the curves rcr versus q and qCT versus r. Both curves can be displayed on a single diagram (Fig. 3, the \"type diagram\"), as OPV (rcr versus q) and OPH (gc, versus r). Both curves have a common branch, OP, which begins at the origin and terminates at the branch point, P. Beyond P, the curves are horizontal and vertical straight lines. These curves divide the q-r plane into regions according to type. For a given combination-of q and r, the range of b must lie within the intermediate cZ-values as shown; otherwise double rockers are obtained, assuming assembly to be possible. Analogous to the planar case, the g-coordinate of P, qP, is that value of q at which the g-crank circle intersects the r-crank axis, ?,-2 = P2 + g\"- tan ! f . (3) By similar reasoning, the r-coordinate of P, rP, is given by >>2 = p2 + / 2 tan2f. (4) These equations can also be derived by noting that (for these values) the skew-torus cross section, which generally has two circuits, Fig. 4(a), has a single circuit, with a double point on the Z-axis, Fig. 4(6). Several conclusions can be drawn from the type diagram, Fig. 3. \u2022 For a plane four-bar linkage, OP is a straight line inclined 45 deg to the g-axis and rP = qP = p, where p now is the length of the fixed link. The diagram may be regarded as a graphical interpretation associated with Grashof's rule. \u2022 For a simply skewed skew four-bar mechanism (zero offsets: / = g = 0\u2014the term is from Harrisberger), the determination of type is identical to that of a planar four-bar linkage with the same link lengths. \u2022 For a skew four-bar mechanism with equal offsets ( / = g) OP is a straight line, inclined 45 deg to the coordinate axis", + " \u2022 It is also possible to compute the dimensions of an equivalent plane four-bar linkage from the values (/, (i = 1,2,3,4), and to apply Cirashof's rule directly to the equivalent, linkage. The proportions of the equivalent linkage can be obtained by transforming the (/-values in accordance with the rules for a plane four-bar linkage; sec the Appendix. Examples The procedure for the determination of type is as follows: 1 Select values for the relative orientation (p,f) of the cranks, and for the offsets (f,g). 2 Draw the type diagram, Fig. 3, using equations (3) and (4) for the coordinates of P, noting the horizontal and vertical branches, and that the branch OP can be reasonably approximated by a straight lino if f and p are small, or if the f/g ratio is close to unity. 3 Compute the intermediate ranges (d2,d3) for am- given point on the type diagram. These determine the range of b. Example 1 Suppose the relative orientation of the cranks is p = 2 in , f = 60 deg, and that the offsets are / = 1, g = 4. Investigate the linkage type when Case A: q = 1 in , r = 4 in" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003415_1.4759127-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003415_1.4759127-Figure1-1.png", + "caption": "FIG. 1. (a) Exploded view of the device, showing the droplet, micro-grippers, silicon substrate, and piezoelectric transducer. (b) Image of particles collected within the micro-grippers using an actuation frequency of 1.480 MHz. Images showing (c) the focal plane level with the top of the grippers and (d) the substrate (base of the grippers), which are clearly distinguishable. Snapshots over time showing the dispersion of particles from (e) a stacked arrangement to ((f) and (g)) a dispersed arrangement after actuation is turned off. This occurs in the focal plane, which coincides with the substrate.", + "texts": [ + "17 A common feature of all of these examples of ultrasonic particle manipulation is that the ultrasonic pressure fields were dictated by the geometric shape of the fluid; either the shape prescribed by the enclosure or, in the case of the droplet, by the liquid-air interface. This letter reports a new method to minimize the dependence of the particle forces on the geometric shape and volume of the fluid. We have developed ultrasonic microgrippers, microfabricated cylindrical structures actuated through piezoelectric elements [Fig. 1(a)], to establish a localised pressure field within the fluid volume. Experiments conducted with fluid droplets, the volume and placement of which were intentionally varied, demonstrated that in the presence of ultrasonic actuation, the grippers could control the resonant frequencies and the position of the particle clumps reliably and consistently for all of the tested cases. Hence, through a combination of microfabricated features and ultrasonic actuation, the localised pressure fields and hence the particle manipulation can gain an independence from the global geometry of the fluid", + " The specific case presented here finds applications in the concentration of particles in small droplet volumes for subsequent detection, either optically,18 using Raman spectroscopy19 or using a biological sensor.20 For all these applications, accurate control of the clump locations and resonant frequencies ensures effectiveness and ease of operation. The ultrasonic micro-grippers used in the experiments consisted of 100 lm deep, segmented cylindrical structures. The inner and outer diameters were 1000 lm and 1220 lm, respectively [Fig. 1(a)]. The grippers were etched into a (100) oriented Si wafer using the Bosch process and the wafer was diced into 10 10 mm2 chips for subsequent testing. Each chip was affixed onto a piezoelectric piece (Pz26, Ferroperm, Denmark) to enable actuation. Prior to assembly, the lower electrode of the piezoelectric was separated into two regions (by a cut of approximately 20 lm depth using a wafer saw) such that a strip of electrode runs down one side of the piezoelectric. Wires were connected to the electrodes using conductive silver paint, the active signal being connected to the strip electrode, and all other parts of the electrodes being grounded.12 Finally, the silicon chip-piezo assembly [Fig. 1(a)] was affixed to a glass microscope slide (not shown) to allow easy mounting under a microscope (Olympus ZSH, Australia) fitted with a CCD camera (Hitachi HV-D30, USA) connected directly to a PC for image capture. For the experiments, a 1.00 ll droplet of distilled water containing a 2% concentration of 4.47 lm diameter silica spheres was introduced manually via a micropipette above the micro-grippers. The device was then actuated with a drive signal, produced using a signal generator (Stanford Research DS345) connected to a power amplifier (Amplifier Research 25A250A). It was observed experimentally that collection of the particles within the centre of the gripper occurred at a frequency of 1.480 MHz [Fig. 1(b)]. The height of the collected particles was determined by altering the focal plane of the microscope. Figs. 1(c) and 1(d) show the relatively large focal difference between particles on the top surface of the a)Author to whom correspondence should be addressed. Electronic mail: adrian.neild@monash.edu. 0003-6951/2012/101(16)/163504/4/$30.00 VC 2012 American Institute of Physics101, 163504-1 Downloaded 12 Jun 2013 to 141.161.91.14. This article is copyrighted as indicated in the abstract. Reuse of AIP content is subject to the terms at: http://apl", + "org/about/rights_and_permissions micro-grippers and those resting at the bottom of the cavity (prior to actuation), respectively. The piezoelectric was then actuated, leading to the formation of a clump, after termination of the excitation the particle clumps were observed to disperse, Figs. 1(e)\u20131(g), this sequence of images was recorded with the focal plane fixed at the base of the gripper. Initially, once the trapping field was removed, the particles at the centre of the image of the clump are out of focus, whilst those at the periphery of the clump are in focus [Fig. 1(e)]. Subsequently, the particles are observed to slide outwards [Fig. 1(f)] and as they rapidly become focussed this demonstrates that they come to rest on the bottom surface [Fig. 1(g)]. These observations suggest that during ultrasonic actuation the agglomerated particles sit on the substrate, with the force field causing particles to be stacked in layers. In order to demonstrate that the particle collection shown in Fig. 1(b) was due to a local pressure field induced by the micro-grippers and not a global property of the droplet, a series of experiments were performed in which droplets were placed off centre from the grippers. The respective centroids of the droplets, collected particles, and micro-grippers were determined from image analysis; they are as shown by the square, cross, and diamond markers superimposed on the images in Figs. 2(a) and 2(b). The results of twelve of these tests are summarized in the polar plot of Fig", + " 2(d) showing the collection within a 1.05 ll droplet at the same actuation frequency. Since the collection is not finely tuned to the position or volume of the droplet, it is evident that a localised pressure field attributable to the micro-grippers is responsible for the particle collection. To further illustrate the effect of the grippers for particle collection, a series of finite element simulations were per- formed using COMSOL MULTIPHYSICS. An exploded view of the modelled geometry is presented in Fig. 1(a). The acoustics module of COMSOL was applied to the droplet, to solve the Helmholtz equation \u00f0r2 \u00fe k2\u00dep \u00bc 0; (1) where k is the wavenumber and p the acoustic pressure. The Helmholtz equation is widely used for ultrasonic application where the fluid can be considered inviscid and compressible. It determines the pressure field from which the force field on the suspended particles can be found for specified boundary conditions. A model of the system, including the droplet, silicon substrate, and piezoelectric transducer, was constructed", + " 3(a) shows the resultant modelled radiation force potential field when the droplet was symmetrically located over the gripper and the piezoelectric was actuated at 1.57 MHz (the resonant frequency within the grippers, as defined by the finite element model). It can be seen that the micro-grippers localized the U field in the 1 mm diameter region prescribed by the gripper. The particles are expected to cluster in the location of minimum force potential (dark blue) and as such, for this case, with sufficient particle concentration the line shape seen in Fig. 1(b) would be produced. At lower concentrations, the particle cluster will become more circular in shape, conforming to contours in the radiation force potential. When the droplet was offset by 10% of its diameter in the x direction [Fig. 3(b)], a similar (slightly rotated) outcome resulted at 1.56 MHz, with the particles again collected at the centre of the gripper. (It should be noted that this small frequency shift was not observed in the experiments.) From the predicted force potential field patterns and required driving frequencies, we can confirm that the grippers were acting to create a localised field, which lessens the effect of the fluid geometry" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001766_la900185a-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001766_la900185a-Figure5-1.png", + "caption": "Figure 5. Demonstration of (a) the springlike legs of a water strider when standing on the water and (b) the antivibration mechanism.", + "texts": [ + "6N/mand kw=0.3N/m,which indicates that the small water-treading stiffness has a partial contribution from the flexibility of a water strider\u2019s leg. The SEM image of the cross section of a peeled leg, as shown in Figure 4,was found tobehollowwith 4/5 diameter, which not only lessens the weight of the insect and provides more buoyancy but also allows the leg to be more flexible. In fact, each leg is like a spring reacting independently to outgoing forces; its six legs construct a spring matrix, as shown in Figure 5a. The water strider\u2019s legs can flexibly adjust when the water surface is unstable and maintain the balance of its entire body, as demonstrated in Figure 5b. 3-2. Tiny Friction and Large Thrust. For a water strider\u2019s leg, the drag forces against forward flows were found to be much smaller than those against lateral forces. According to fluid dynamics, a drag force resulted from the friction and pressure forces. The diameter of a driving leg is approximately 0.2 mm, 7008 DOI: 10.1021/la900185a Langmuir 2009, 25(12), 7006\u20137009 which is 1/15 the length beneath the free surface. Therefore, the frontal area for the forward flow was much smaller, which resulted in a smaller pressure force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002761_icppw.2012.69-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002761_icppw.2012.69-Figure6-1.png", + "caption": "Figure 6. Two swing trajectories.", + "texts": [ + " For each sensor, the average distance between the five bias vectors is calculated by 1 n2 n\u2211 j1=1 n\u2211 j2=1 \u2225bi,j1 \u2212 bi,j2\u2225 . (11) \u2022 Let gi,j,k denote the k-th g-vector in the j-th data set of the i-th sensor. To evaluate the scaling errors of each sensor, for each sensor, we calculate the RMSE of the norm of the g-vectors after calibrating the offset errors by\u221a\u221a\u221a\u221a 1 mn n\u2211 j=1 m\u2211 k=1 (\u2225gi,j,k \u2212 bi,j\u2225 \u2212 \u2225G\u2225)2. (12) All numerical results of the evaluation can be found in Table. I. According to the results, we can see the highcost IMU are outperformed to the smartphone build-in g-sensors. Fig. 6 illustrates two swing trajectories of a No. 7- iron club calculated by our algorithm from the IMU data collected at a sampling rate of 200Hz. In the figure, the line in lighter color illustrates the backswing trajectory and the darker one illustrates the downswing trajectory. The backswing is the action in which the golf club is moved rearward away from the ball and cocked up until reaching the top position. The downswing is the action in which the golf club is swung down from the top position until reaching the impact position. Fig. 6(a) is the trajectory of a small trial swing, and Fig. 6(b) shows the trajectory of a full swing. Fig. 7 illustrates a typical waveform of the g-sensor readings and gyroscope readings. The sampling rate in this case is 100Hz. Three key frames of a golf swing, including the starting frame, the top of swing and the impact frame, are marked. The photo snapshots at these three key frames, labeled by (a), (b) and (c) respectively, are also provided. The key frames can be identified according to the readings of the g-sensor and gyroscope. Besides the trajectory itself, the speed of the golf club is also known as we calculate the trajectory" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003712_01691864.2015.1023219-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003712_01691864.2015.1023219-Figure2-1.png", + "caption": "Figure 2. Kinematical scheme of first leg A of lower module.", + "texts": [ + " The active leg A of lower module, for example, consists of a little cross of a universal joint of mass m1linked at the frame A1xA1y A 1 z A 1 which has the angular velocity xA 10 \u00bc _uA 10 and the angular acceleration eA10 \u00bc _xA 10 connected at a moving cylinder A2xA2y A 2 z A 2 of length l2, mass m2 and tensor of inertia J\u03022, having a relative rotation around A2zA2 axis with the angle uA 21, so that xA 21 \u00bc _uA 21, e A 21 \u00bc _xA 21. An actuated prismatic joint is as well as a piston of length l3, mass m3 and tensor of inertia J\u03023, linked to the A3xA3y A 3 z A 3 frame, having a relative velocity vA32 \u00bc _k A 32 and acceleration cA32 \u00bc _vA32. Finally, a ball-joint or a spherical joint A4 is attached at the edge of first moving platform, which can be schematized as a circle of radius rp, mass mp and central tensor of inertia J\u0302p (Figure 2). At the central configuration, we also consider that all 12 active sliders are starting from same initial relative position A2A3 \u00bc A 2A 3 \u00bc B2B3 \u00bc B 2B 3 \u00bc ::: \u00bc F2F3 \u00bc F 2F 3 \u00bc l1 and that the angles of orientation of universal joints and spherical joints are given by where \u03b2 is an angle of initial inclination. A dv an ce d R ob ot ic s Since the hybrid manipulator is an assemblage of links and joints, this can be symbolized in a more abstract form known as equivalent graph representation, using the associated graph to represent the topology of the mechanism (Figure 3)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002485_1.4025350-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002485_1.4025350-Figure2-1.png", + "caption": "Fig. 2 Definition of the needle bearing parameters used in the load distribution model", + "texts": [ + " First of all, any power loss taking place along the contact between the cage and the rollers is assumed to be negligible. The rollers (or needles) are considered to remain equally spaced from each other at a spacing angle of / \u00bc 2p=Z where Z is the number of rollers in the bearing. As such the position angle of roller z is defined to be /z \u00bc \u00f0z 1\u00de / (z 2 \u00bd1;Z ) with /1 \u00bc 0. The rollers are assumed to remain parallel to the nominal rotational axis of the planet (i.e., they are not allowed to skew), as well as to each other as illustrated in Fig. 2. Furthermore, in the case of double-row needle bearings, the right and left needles are assumed to line up perfectly with one another. Material parameters common to bearing steels are employed in the model. The rollers are divided into a prescribed number of K axial slices of equal width. The radial force balance equation and overturning moment balance equation for the entire bearing (in two dimensions), as well as the equation of contact for each roller z (z 2 \u00bd1;Z ) are given, respectively, in a generalized form based on Ref", + " [19] 1:24\u00f010\u00de 5Fx w0:89 XZ z\u00bc1 cos /z k0:11 z Xk\u00bckz k\u00bc1 Dz 6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 \u00bc 0 \u00bc Uj (2a) 1:24\u00f010\u00de 5Fy w0:89 XZ z\u00bc1 sin /z k0:11 z Xk\u00bckz k\u00bc1 Dz 6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 \u00bc 0 \u00bc Uj (2b) 1:24\u00f010\u00de 5My w0:89 XZ z\u00bc1 cos /z k0:11 z Xk\u00bckz k\u00bc1 Dz 6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 k 1 2 w ( Xk\u00bckz k\u00bc1 1 2 \u2018 Dz 6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 ) \u00bc 0 \u00bc Uj (2c) 1:24\u00f010\u00de 5Mx w0:89 XZ z\u00bc1 sin /z k0:11 z Xk\u00bckz k\u00bc1 Dz 6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 k 1 2 w ( Xk\u00bckz k\u00bc1 1 2 \u2018 Dz 6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 ) \u00bc 0 \u00bc Uj (2d) dx6 1 2 \u2018hy cos /z \u00fe dy6 1 2 \u2018hx sin /z Cd 2 2 Dz6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck max \u00bc 0 \u00bc Uj; z 2 \u00bd1;Z (2e) 121007-2 / Vol. 135, DECEMBER 2013 Transactions of the ASME Downloaded From: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jmdedb/927630/ on 02/01/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Referring to Fig. 2, dx and dy are the radial, and hx and hy are the angular deflections of the bearing, \u2018 is the roller length, Cd is the bearing diametral clearance, Dz is the mean deflection of roller z, k is the index of the axial roller slices (k 2 \u00bd1;K ), w is the width of each slice, /z is the azimuth angle of roller z, ck is the magnitude of the roller crown drop at slice k, Fx, and Fy are the bearing radial loads, Mx and My are the bearing moment loads, and kz is the number of slices of roller z that are in contact", + "org/about-asme/terms-of-use Qz \u00bc w0:89 1:24\u00f010\u00de 5k0:11 z Xk\u00bckz k\u00bc1 Dz6 1 2 k 1 2 w hy cos /z \u00fe hx sin /z ck 1:11 ; z 2 \u00bd1;Z (4b) 3.1 Mixed EHL Formulation. Assuming that relative sliding at the contacts of each roller with the inner race (planet pin) and outer race (planet gear bore) are negligible, the velocity of a roller-race (inner or outer) contact can be considered to represent a pure rolling condition. The rolling mechanical power loss of a bearing can be determined by employing an EHL formulation. The contacts between a roller slice in Fig. 2 and the inner and outer races are modeled as line contacts such that the pressurized hydrodynamic fluid flow is governed by the one-dimensional transient Reynolds equation as @ @x qh3 12g @p @x \u00bc @\u00f0urqh\u00de @x \u00fe @\u00f0qh\u00de @t (5a) where p\u00f0x; t\u00de, h\u00f0x; t\u00de, g\u00f0x; t\u00de, and q\u00f0x; t\u00de are the normal pressure, film thickness, lubricant viscosity, and density distributions along the rolling direction x at the time instant t. The rolling velocity ur is given by the rolling element angular velocity as ur \u00bc \u00f01=4\u00dedmxb\u00f01 d2 r =d2 m\u00de with dr representing the diameter of the roller slice", + " Under these typical conditions, radial bearing force, Fr , and the bearing moment, M, reduce to Fr \u00bc 2W\u00f0t\u00dem \u00bc 2TPB rps (14) M \u00bc 2W\u00f0a\u00dem rpp \u00bc 2TPBrpp tan b rps (15) where TPB is the sun gear torque carried by a planet branch (i.e., TPB \u00bc TS=n where TS is the external torque applied to the sun gear and n is the number of planet branches in the gear set), b is the helix angle, rpp is the planet pitch radius, and rps is the sun pitch radius. While the formulation proposed in Sec. 2 applies to any arbitrary combination of Fr and M, Eqs. (14) and (15) show that Fr and M of a planet bearing are related to each other as both are defined by parameters b, rpp, and rps. It is noted here that in Fig. 2 Fr is in the direction of any arbitrary combination of Fx and Fy; however, the moment applied by this planet loading is a couple applied about the same direction. Meaning if Fr \u00bc Fx, then M \u00bc Mx. Meanwhile, the planet bearing speed, Xb, represents the planet gear speed with respect to the carrier. In this example system, rpp \u00bc 24:2 mm, rps \u00bc 67:9 mm and b \u00bc 13:1 deg. Predicted bearing load distributions of the example bearing are shown in Fig. 4 at load levels of TPB \u00bc 42, 100, and 333 Nm. Here, the zero degree position represents the direction the radial force applied as shown in Fig. 2 where Fr \u00bc Fx. Roller #1 is placed at this position such that /1 \u00bc 0. Rollers on the opposite side of the bearing from this zero degree position remain unloaded. Figure 4 shows equal and uniform load distributions along the length of the roller at this zero degree position, while the load distributions of other rollers become more skewed with an increase in the needle position angle, /z, highlighting the effect of M on the bearing. Since roller #1 at /1 \u00bc 0 is positioned orthogonal to the loading produced by M, its load distribution is not influenced (skewed) by M" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002769_j.measurement.2011.07.009-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002769_j.measurement.2011.07.009-Figure5-1.png", + "caption": "Fig. 5. Schematic diagram of the camera stations.", + "texts": [ + " The procedure based on photogrammetry technology is divided into the following stages: (1) Paste un-coded points and coded points on or around the object, select suitable scale bars for the measurement. Generally, the scale bars fit 1:1 to the object to be measured (e.g. the longest extension of the object is approximately 1 m and the length of the scale bars is approximately 1 m as well). To measure a simple flat object, you may need eight images, four calibrating images taken from the top and four images recorded laterally at an angle of approximately 90 . (2) Hold the camera in half of object height, as shown in Fig. 5, walk around the object and record images at a distance of 45 from each other. In each location, two photos (horizontally and vertically) must be taken as original 2D data. Then, record the same number of images in each level. (3) The images are processed with sub-pixel extraction technology based on least squares fitting method. And then the exact geometric centers of codedpoints and un-coded points are obtained. The coordinates and code information of the above coded points can be solved using collinear equation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000155_j.triboint.2007.02.018-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000155_j.triboint.2007.02.018-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of a pair of meshing helical gears.", + "texts": [ + " In this paper, full numerical TEHL solution of two tapered rollers in opposite orientation is obtained by using a multilevel solver. The TEHL solution of cylindrical rollers [10] is achieved again as a special case using the program of tapered rollers, and then comparisons are made between our solution of the cylindrical rollers and the solution of the finite line contact obtained by Liu and Yang [10]. The influences of some parameters, such as the angle of rollers and the velocity of rotating, on the lubrication are analysed. The meshing teeth of a pair of helical gears are shown in Fig. 1. K1K2 is the meshing line, N1N2N3N4 the meshing plane, and B1B2B3B4 the actual meshing plane. b1 and b2 are the spiral angles between K1K2 and two axes (for helical gears, b1 b2). At the moment of meshing shown in Fig. 1, the meshing teeth of the helical gears can be transformed approximately to two tapered rollers (a and b) in opposite orientation as shown in Fig. 2. The axes of rollers a and b shown in Fig. 2 correspond to lines N4N3 and N1N2, respectively in Fig. 1, and the contact line of two tapered ARTICLE IN PRESS P. Yang, P. Yang / Tribology International 40 (2007) 1627\u20131637 1629 rollers shown in Fig. 2 corresponds to line K1K2 in Fig. 1. In Fig. 2, ra and rb are the section radii on the middle section of tapered rollers a and b, respectively, ral and rbl are the section radii at the large ends of tapered rollers, respectively, similarly, ras and rbs are the section radii at the small ends. ba and bb are the deflective angles between the common generatrix and both axes of the two rollers, respectively. Geometric relationships between the helical gears shown in Fig. 1 and the tapered rollers shown in Fig. 2 are as follows: rbl \u00bc K1N1; rbs \u00bc K2N2; ral \u00bc K2N3; ras \u00bc K1N4, ba \u00bc b1; bb \u00bc b2. A Cartesian coordinate system as shown in Fig. 2 is established for the oil film, where the x-axis is perpendicular to the sheet of paper and points to readers. The coordinate systems for solids a and b are similar to that for the film, that is, the x- and y-axes are the same as those shown in Fig. 2, and, za and zb coordinates are used for solids a and b, respectively, with the same direction as z, however, with different origins" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002184_cp.2012.0275-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002184_cp.2012.0275-Figure2-1.png", + "caption": "Figure 2: Inverse cosine airgap magnet.", + "texts": [ + " As stated above, the magnet shape can be optimized to achieve sinusoidal airgap flux density. The inverse cosine airgap length and sinusoidal shape magnets can be easily determined by analytical method, which thus will be employed in this paper. The airgap length is varied with the angle from the middle of magnets and, thus, the magnet shape produces sinusoidal flux distribution. The airgap length can be expressed as [1, 8]: )cos( )( p gd g l l (1) where p is pole pitch, lgd is the airgap length along d-axis. The variations of magnet shape and airgap length are shown in Figure 2. The finite element predicted airgap field distribution without slot-opening is shown in Figure 3, which shows that the airgap flux distribution is non-sinusoidal due to the influence of the pole arc to pole pitch ratio, but the harmonics of airgap flux density are still very small. Another method of sinusoidally shaped arc magnets, as shown in Figure 4 (a), has been proposed to reduce the cogging torque and make back-EMF waveform sinusoidal distribution [14]. The magnet edges are assumed to be radial and the maximum thickness of the arc magnet is 3mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001497_j.matdes.2010.02.029-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001497_j.matdes.2010.02.029-Figure8-1.png", + "caption": "Fig. 8. Stress pattern obtained using ANSYS.", + "texts": [ + " This enables user to apply different process parameters for single layer (Fig. 7). Developed methodology can reduce weight of parts by applying low energy density and faster prototyping by applying higher laser velocity and scan spacing in non-critical regions obtained by CAE analysis. Further the method can be used in design and fabrication of functionally graded materials. A case study of crane hook is presented here to describe implementation procedure. Typical stress distribution for crane hook is obtained using FEM analysis and presented in Fig. 8. Cross-sectional images of stress levels are retrieved at each section along the Z direction similar to the build direction in RP process. These cross-sectional images of stress patterns for each layer similar to RP process are stored as jpg images through ANSYS Programming Design Logic (APDL) of ANSYS software. Procedure for storing stress patterns through APDL is presented in Fig. 9. 2D cross sections of crane hook in XY plane along with stress values have been stored as jpg image (Fig. 10). Further stress ranges can be chosen as per the requirements or number of regions one wishes to have" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002326_robio.2010.5723563-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002326_robio.2010.5723563-Figure7-1.png", + "caption": "Fig. 7. Shape primitive: tapered square pillar with a hole and a base", + "texts": [ + " In the same manner as for basic grasping configurations 1, 2, and 3, \u03b8 is finited as \u03b8 = \u03c0 2 k (k \u2208 Z). 2) With a Hole and a Base: The tapered square pillar with a hole and a base is specified by eight parameters: outside length of the short side ls outx, inside length of the short side ls inx, outside length of the long side ls outy , inside length of the long side ls iny , height ls out, depth ls in, taper angle of the short side \u03b8s x, and taper angle of the long side \u03b8s y . The same coordinate system for the shape primitive is used as that for the filled tapered square pillar (Fig. 7(i)). A basic grasping configuration for grasping the brim of this shape primitive is added to the four configurations shown in the filled one (Fig. 7(ii)). For basic grasping configuration 5, the approach direction is determined along the taper direction and the grasping points are represented by the angle about the z axis, \u03b8, the distance from the x or y axis, r, and the depth from the edge h. In the same manner as for the other basic grasping configuration, \u03b8 is finited as \u03b8 = \u03c0 2 k (k \u2208 Z). 3) With a Hole: The tapered square pillar with a hole is specified by seven parameters (Fig. 8(i)). The coordinate system of the shape primitive is the same as that for a filled pillar and a pillar with a hole and a base (Fig. 8(i)). Similarly as for the cylinder primitive, a basic grasping configuration for grasping the brim from the underside of the shape primitive (Fig. 8(ii)) is added to the five configurations shown for the tapered square pillar with a hole and a base (Fig. 7(ii)). Basic grasping configuration 6 is parameterized by the same parameters as the basic grasping configuration 5, but the approach direction is reversed. It is possible to model an object in detail using the small shape primitives. However, many shape primitives reduce the abstraction effectiveness since they increase the number of grasping configurations applicable to an object. Therefore, we aim to abstract a grasping object using the minimum number of shape primitives. At first, a shape primitive is prepared that can wrap the target object" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure1-1.png", + "caption": "Fig. 1. Basic rack profile and design parameters.", + "texts": [ + " As early as 1961, Andreev [1] applied the envelope method\u2014originally introduced in the eighteenth century by Euler and described in 1956 by Litvin [2]\u2014to the generation of screw compressor rotors and their tools. Litvin and Feng [3] used singularity and tooth contact analysis (TCA) to investigate the influence of 0094-114X/$ - see front matter 2008 Elsevier Ltd. All rights reserved. doi:10.1016/j.mechmachtheory.2008.02.013 * Corresponding author. Tel.: +886 5 2720411; fax: +886 5 2720589. E-mail address: imezhf@ccu.edu.tw (Z.-H. Fong). Nomenclature b pitch helix angle of rotor d distance from center line to point A (Fig. 1) e given tolerance / rotation angle c included angle of circular arc EF g upper parametric limit of trochoid CD gV volumetric efficiency gI isentropic indicated efficiency j normal equidistance of trochoid CD k pitch lead angle of rotor l slanted angle of ellipse FG h curve parameter of rack h1, h2 lower and upper parameter limits of ellipse FG q1 radius of circular arc AB q2 radius of circular arc HI r1 upper limit of torque ratio of two rotors s normal equidistance of trochoid DE x rotation speed -i weighted factors, i = 1,2 n penalty parameter f upper parametric limit of trochoid DE Ac center distance between two rotors C increasing coefficient d length of straight line IJ ea major radius of ellipse FG eb minor radius of ellipse FG ex, ey coordinates of center point of ellipse FG F o penalized objective function fm equation of meshing fo objective function g \u00f0g1; . . . ; gng \u00de set of ng implicit constraints ha addendum of rack hd dedendum of rack k iteration number L length of rotor screw part l1i, l2i feasible region of ith design variable m number of design variables m12 ratio of rotation speeds \u00bc x1=x2 gg number of constraints ni unit normal vector in coordinate system Si P penalty function p distance from point CI to yh-axis (Fig. 1) rd inner radius of rotor ro outer radius of rotor rp pitch radius of rotor rT ratio of average torque on female rotor to average torque on male rotor ri\u00f0xi; yi; zi\u00de position vector in coordinate system Si Si coordinate system i s length of straight line GH t length of straight line BC U i condition of non-undercutting, i = x,y u pressure angle in high-pressure side of rack v pressure angle in low-pressure side of rack Vi sliding velocity in coordinate system Si W transverse circular pitch of rack w design variables z tooth number of rotor Subscripts 1 male rotor 2 female rotor misalignment on the backlash between the surfaces", + " Practical experiments were also conducted to verify this result. An explicitly defined rack is disclosed here to generate a pair of conjugated rotor profiles for twin-screw compressors. The fundamental idea is derived from the rack-generated profile invented by Stosic [5]. Each compound curve has at least one control parameter on the rack profile, which makes the rack both flexible and instinctively adjustable. The tooth profile of the basic rack depends on the pitch W and the total tooth height (ha + hd). As shown in Fig. 1, the addendum ha and the dedendum hd can be determined by the tooth number of the rotors z1 and z2, the center distance Ac, and the outer radii of the male and female rotors ro1 and ro2 as given in the following equations: rp1 \u00bc z1Ac=\u00f0z1 \u00fe z2\u00de \u00f01\u00de rp2 \u00bc z2Ac=\u00f0z1 \u00fe z2\u00de \u00f02\u00de ha \u00bc rp1 rd1 \u00bc ro2 rp2 \u00f03\u00de hd \u00bc ro1 rp1 \u00bc rp2 rd2 \u00f04\u00de W \u00bc 2p z1 rp1 \u00bc 2p z2 rp2 \u00f05\u00de where rp1 and rp2 are the pitch radii of the male and female rotors, and rd1 and rd2 are the inner radii of the male and female rotors, respectively. As also shown in Fig. 1, the coordinate system Sh is attached to the rack cutter, while xh-axis is collinear with the pitch line of the basic rack. The position vector rh(h) = [xh(h),yh(h)] of each rack segment is listed in Table 1. The unit normal vector of the rack nh(h) = [nxh(h),nyh(h)] can be derived from the following equation: nh\u00f0h\u00de \u00bc k ohrh\u00f0h\u00de jk ohrh\u00f0h\u00dej \u00f06\u00de The normal-equidistant trochoids CD and DE are the specific and significant segments which are deserved to explore. As the generating principle of normal-equidistant trochoids of the proposed rack shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002063_j.triboint.2010.11.020-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002063_j.triboint.2010.11.020-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of piston-", + "texts": [ + " The piston ring motion equation is as follows [42]: Mr d2h dt2 \u00bc Fp\u00feFf \u00feFi\u00feFs\u00feFasp \u00f01\u00de where Mr is the mass of each ring, h the oil film thickness, t the time, Fp the pressure force, Ff the friction force, Fi the inertia force, Fs the radial hydrodynamic force and Fasp the radial contact force on the ring face. The gas flow equations through the piston-ring pack crevice are as follows [43]: m02 P02 dP2 dt \u00bcm12 m23 \u00f02\u00de m03 P03 dP3 dt \u00bcm13 \u00fem23 m34 m35 \u00f03\u00de m04 P04 dP4 dt \u00bcm34 m45 \u00f04\u00de m05 P05 dP5 dt \u00bcm35 \u00fem45 m56 m57 \u00f05\u00de m06 P06 dP6 dt \u00bcm56 \u00fem76 m6c2 m6c1 \u00f06\u00de m07 P07 dP7 dt \u00bcm57 m76 m7c1 \u00f07\u00de where moi is the initial mass of i volume, mij the mass flow rate into j volume from i volume, Pi the pressure of i volume and Poi the initial pressure of i volume. As shown in Fig. 1, 1 indicates the top land clearance, 2 the volume behind the top ring, 3 the volume of second land clearance, 4 the volume behind the second ring, 5 the volume of the third land clearance, 6 the volume behind the oil ring and 7 the volume between the oil ring rails. c1 and c2 indicate the crankcase reached through the piston skirt clearance and oil drain holes, respectively. The gas mass flow rates between adjacent lands are calculated using the orifice flow equation such that m \u00bc CdrAcZ \u00f08\u00de where Cd is the discharge flow coefficient [28], r the gas density, A the flow area associated with the orifice, c the speed of sound and Z the compressibility factor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003260_s11665-012-0445-3-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003260_s11665-012-0445-3-Figure12-1.png", + "caption": "Fig. 12 An example of typical results pertaining to the spatial distribution and temporal evolution of temperature in the weld region at relative welding times of: (a) 10 s, (b) 20 s, (c) 30 s, and (d) 40 s", + "texts": [ + " For example, the results pertaining to the spatial distribution of the equivalent plastic strain and the residual von Mises stress in the weldment could be potentially quite important relative to the overall mechanical performance of the GMAW joint. However, these aspects of GMAW modeling are beyond the scope of this study. In sharp contrast, detailed results pertaining to the spatial distribution and temporal evolution of the temperature within the FZ and the HAZ are the key input to the computational analysis (presented in this section) dealing with the prediction of the material microstructure and phase volume fractions, within the weld region. 4.1.1 Temporal Evolution of the Weldment Temperature Field. Figure 12(a) to (d) shows typical results pertaining to the temporal evolution of the temperature field within the weld region. The results displayed in Fig. 12(a), (b), (c), and (d) are obtained at relative welding times of 0, 10, 20, and 30 s, respectively. To improve clarity, regions of the weldment with a temperature lower than 400 K are denoted using light gray. Examination of the results displayed in Fig. 12(a) to (d) reveals that: (a) filling of the groove with the filler-material gives rise to an abrupt increase in temperature in the region next to the workpiece/weld interface; (b) as welding proceeds, natural convection and radiation to the surroundings, together with conduction through the adjacent material region, cause the previously deposited filler material to cool; and (c) by monitoring the expansion of the 400 K temperature contour over the workpiece top surface, one can visualize the propagation of the thermal-conduction wave within the workpiece" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001677_s00170-010-2705-4-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001677_s00170-010-2705-4-Figure8-1.png", + "caption": "Fig. 8 Forces in the contact plane\u20143D dynamic case", + "texts": [ + " Friction forces due to relative translational and rotational motion links are always parallel or perpendicular to the Z axis. A key difference between the current analysis and previous work [1, 6, 30, 31] is that tangential forces at contact points are considered in the spatial analysis due to the permitted rotation of the outer link about its axis. The forces at contact points, i.e., normal force, friction due to sliding and friction due to rotation, are considered to be orthogonal to each other. Figure 8 describes the forces on the contact plane. Only Z and Zp axes are specified as the reference frames due to the possible rotation of the contact plane about axis Zp, and thus indicating that axis Xp does not always coincide with the contact plane. As shown in Fig. 8, the effective acceleration on the contact plane due to XE and \u0178E is XE cos a \u00fe YE sin a , where XE and YE are the acceleration components of the end-effector and are considered to be similar to those of the Cartesian manipulator along the Xp and Yp directions, respectively. By D\u2019Alembert\u2019s principle, the contact forces may be written as fa \u00bc 1 2m \u00bd F2 cos b F3 sin b\u00f0 \u00de sin q m cos q\u00f0 \u00de F1 cos q\u00fe m sin q\u00f0 \u00de mm :: X p cos a \u00fe :: Yp sin a mg m :: Zp m :: Ze \u00f015\u00de and fb \u00bc 1 2m \u00bd F2 cos b F3 sin b\u00f0 \u00de sin q\u00fem cos q\u00f0 \u00de F1 cos q m sin q\u00f0 \u00de \u00fe mm :: X p cos a \u00fe :: Yp sin a mg mZp mZe \u00f016\u00de Since sin\u03b8\u22480 and cos\u03b8\u22481 when \u03b8 is very small, Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002650_s12540-011-0223-z-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002650_s12540-011-0223-z-Figure2-1.png", + "caption": "Fig. 2. Pin-on-disk test: (a) test system and (b) dimensions of the disk and the pin.", + "texts": [ + " During the tests, a data logging system continuously recorded the test time (sliding distance), the friction coefficient, the torque and the normal force. The data were written to a file for further processing. The disks had a diameter of 54.9 mm and a height of 6 mm. The disk was machined from H13 die steel of which the hardness was approximately 200 HV. The pin was shaped so that the disks could be firmly mounted, and the disks had a circular contact area with a diameter of 2.5 mm. The average surface roughness of both the pin and the disk after NC-machining was 0.3 \u00b5m to 0.4 \u00b5m. Figure 2 shows the dimensions of the disk and the pin. The contact surfaces of the pin were a melted layer with a thickness of approximately 0.2 mm formulated with the Fe-Ni and the Fe-Cr powders using an energy input of 11.38 J/mm2 (a laser power of 200 W and a scan rate of 219.73 mm/s). To measure the wear rate and friction, a wear test was carried out using a POD testing system under a varying load (20 N, 40 N and 60 N) at a sliding speed of 200 rpm. Prior to the test, the pin and the disk were cleaned in an ultrasonic cleaner" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001400_00423110903126478-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001400_00423110903126478-Figure2-1.png", + "caption": "Figure 2. A convex set with exemplary slip velocity and the friction force.", + "texts": [ + " The mathematical description of this principle has the form of the following variational inequality: Vt s \u00b7 T \u2265 Vt s \u00b7 T\u0303 \u2200 T\u0303 \u2208 , (4) where T\u0303 \u2208 R2 is the variational friction force. The variational inequality (4) uniquely assigns the vector T of the set to the slip vector Vs and vice versa. The ambiguity arises when Vs = 0 because then any force of the set satisfies the inequality (4). For any friction force belonging to the interior of the set , T \u2208 Int the inequality (4) assigns the slip equal to Vs = 0 (sticking). In Figure 2, showing a smooth convex set , such a force T is marked that dissipates energy with the maximum power when the slip is Vs. The geometrical determination of the force T is evident from the drawing. The angle \u03b2 is called the slip direction angle. The angle \u03b1 is the friction force direction angle. It is easy to prove by elementary mathematical manipulations that the geometrical construction shown in Figure 2 is indeed valid. D ow nl oa de d by [ O tte rb ei n U ni ve rs ity ] at 0 3: 30 2 1 A pr il 20 13 In case of the orthotropic elliptical set := { T \u2208 R2, (T) := ( Tx T0x )2 + ( Ty T0y )2 \u2212 1 \u2264 0 } , the angle \u03b1 is given by the expression tan \u03b1 = T 2 0y T 2 0x tan \u03b2, (5) and the break-out force T 0 at the angle \u03b1 has the value T0 = T0xT0y\u221a T 2 0x sin2 \u03b1 + T 2 0y cos2 \u03b1 , (6) where T 0x and T 0y are the break-out forces in the X- and Y -directions. In case of the isotropic friction T 0x =T 0y =T 0, the friction force direction angle is equal to the slip direction angle, \u03b1 =\u03b2", + " The axis of the element belongs to a common plane of contacting bodies and each end of the element is connected with a different body of the friction pair. It follows from the maximum principle that for isotropic friction, the axis of the element is co-linear with the slip. When there is sticking, the axis of the element is co-linear with the tangential load of contacting bodies. For anisotropic friction, the force is determined by the principle of the maximum power of dissipation by the angle \u03b1, as shown in Figure 2. It has been shown in [9] that the friction force of the element is described by the following differential equation: T\u0307 = \u23a7\u23aa\u23a8 \u23aa\u23a9 K1(\u03bd\u0307 \u2212 Z\u0307) if |T | < T0, \u2212[\u2212K1(\u03bd\u0307 \u2212 Z\u0307)]+ if T = +T0, [K1(\u03bd\u0307 \u2212 Z\u0307)]+ if T = \u2212T0, (7) where [u]+ = { u if u \u2265 0, 0 if u < 0 . This description needs to be supplemented by the determination of the friction force direction angle \u03b1 according to the above described principle. An advantage of the description by the differential equation is the possibility of the direct application of the model of friction to multi-body computer codes, oriented for vehicle system dynamics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003394_2013-01-0633-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003394_2013-01-0633-Figure7-1.png", + "caption": "Figure 7. Test Configuration of Tire with SoMat 2000 Field Computer", + "texts": [ + " In this study, the sampling frequency was set to be 1000 Hz which enables recording of at least 50 data points per tire revolution at rolling speed of 100 km/h. However, considering the minimum number of data acquisition values on the length of the contact patch, the speed used in this study was limited to 30 km/h. The measurement range of SoMat strain gauge module is from \u22125000 microstrain to 5000 microstrain, which is suitable for the normal working range of the 175/505 R13 tire used herein. The SoMat 2000 Field Computer was attached to the wheel and balanced as shown in Figure 7. Measured Strain Data Preliminary measurements of tire dynamic strain on the inner tread were performed under tire steady state straight line rolling and cornering. Before carrying out the test, the measuring system was calibrated and zeroed. The resolution of the strain measurement is 0.001 microstrain. Three channels of test data including two axial direction strains and one circumferential direction strain are taken as example and visualized in the software SoMat InField as shown in Figure 8 and Figure 9" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002780_j.apm.2011.09.043-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002780_j.apm.2011.09.043-Figure2-1.png", + "caption": "Fig. 2. Model of force diagram for pinion and gear.", + "texts": [ + " Section 2 derives dynamic models for the gear-bearing system with a nonlinear suspension effect, turbulent flow assumption, strongly nonlinear gear mesh force and strongly nonlinear oil-film force. Section 3 describes the techniques used in this study to analyze the dynamic response of the gear-bearing system. Section 4 presents the numerical analysis results obtained for the behavior of the gear-bearing system under various operational conditions. Finally, Section 5 presents some brief conclusions. Fig. 1 presents a schematic illustration of the dynamic model considered in the present analysis. To simplify the whole dynamic system, the gyroscopic effect of rotor dynamics is neglected. Fig. 2 presents a schematic illustration of the dynamic model considered between gear and pinion. Og and Op are the center of gravity of the gear and pinion respectively; O1 and O2 are the geometric centers of the bearing 1 and bearing 2, respectively; Oj1 and Oj2 are the geometric centers of the journal 1 and journal 2, respectively; m1 is the mass of the bearing housing for bearing 1 and m2 is the mass of the bearing housing for bearing 2; mp is the mass of the pinion and mg is the mass of the gear; Kp1 and Kp2 are the stiffness coefficients of the shafts; K11, K12, K21 and K22 are the stiffness coefficients of the springs supporting the two bearing housings for bearing 1 and bearing 2; C1 and C2 are the damping coefficients of the supported structure for bearing 1 and bearing 2, respectively; Km is the stiffness coefficient of the gear mesh, Cm is the damping coefficient of the gear mesh, e is the static transmission error and varies as a function of time" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002160_2011-01-1548-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002160_2011-01-1548-Figure3-1.png", + "caption": "Figure 3. Shaft bearing assembly: (a) physical structure, (b) beam finite element representation.", + "texts": [ + " Note that the mesh formulation established above is in the mesh model coordinate system denoted by Sm Therefore, it should be transformed into pinion or gear coordinate systems when assimilated into the dynamic model. Support stiffness matrices for the pinion [kp] and gear [kg] can be obtained computationally, analytically or experimentally. In this study, the FE approach utilizing beam elements is applied to extract the flexibility of shat-bearing assemblies. This proposed approach is applied to model the pinion/gear overhung position as shown in Figure 3. Figure SAE Int. J. Passeng. Cars - Mech. Syst. | Volume 4 | Issue 2 1041 3(a) is the physical structure and Figure 3(b) is the corresponding FE representation. In the shaft-bearing assembly, each rolling element bearing is represented by a set of time-varying stiffness matrices. Pinion/gear shaft is modeled as beam elements, and each node has 6-DOF. The relation between the applied load and deflection at the reference point will give the support stiffness matrix. In this investigation, six load cases are designed. In each load case, one column of a 6\u00d76 identity matrix is taken as the excitation force vector {Fi}, i = 1,2,\u20266 exerted on the reference point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001098_1077546309106143-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001098_1077546309106143-Figure1-1.png", + "caption": "Figure 1. Schematic of the power assist system.", + "texts": [ + " During the sliding mode, although the equivalent control of u 1 rejects the disturbance, it is worth noting that the nominal control u0 is affected by the saturation. Suppose that the actuator output is saturated at the positive value, Ulim. Actually it follows that u0 Ulim d, since u0 u 1eq u0 d Ulim. Therefore, the nominal control should be designed taking this constraint into account, which is beyond the scope of the present paper. In order to investigate basic problems concerning electric power assist systems, we developed a simple manipulator with a single degree of freedom as shown in Figure 1. Eliminating the coil inductance in the motor, the mathematical model is given by Je n2 Kt K R Mgl sin Th nKt R u Td (19) at OhioLink on March 15, 2015jvc.sagepub.comDownloaded from where: Je: equivalent inertia including all rotating parts, : arm angle, M: equivalent mass of load and arm, g: gravitational constant, l: arm length, Th: operational torque by operator, Ta: assist torque, Td : disturbance torque, n: gear ratio, Kt : torque constant of motor, K : back emf constant, R: resistance, u: input voltage Using the state vector, x [ ]T, the state space model of the plan is represented by x1 x2 0 1 0 n2 Kt K R Je x1 x2 0 1 Je Mgl sin x1 0 1 Je Th 0 nKt R Je u 0 1 Je Td Ax b0 Mgl sin x1 b0Th bu b0Td (20) Considering a desired assist rate, maneuverability and so on, a desired closed-loop dynamics can be designed as follows (Yokoyama et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000598_cjece.2008.4721636-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000598_cjece.2008.4721636-Figure4-1.png", + "caption": "Figure 4 : Application de la modulation vectorielle.", + "texts": [ + " Quand le re\u0301gime d\u2019e\u0301quilibre est atteint, comme c\u2019est montre\u0301 sur la figure 3, on obtient la situation suivante : \u03a6s(k) = \u03a6 6=s (k), \u03b4(k) = \u03b4 6=(k). (10) Par conse\u0301quent, on a la relation : \u2206\u03a6s(k, k + 1) = \u03a66= s (k) \u00b7 \u03c9r \u00b7 Te. (11) Pour de\u0301terminer les e\u0301tats de l\u2019onduleur et leurs dure\u0301es d\u2019application, nous avons utilise\u0301 une technique de modulation vectorielle. Ainsi, une combinaison de trois vecteurs, deux actifs et un nul, est applique\u0301e durant chaque pe\u0301riode d\u2019e\u0301chantillonnage pour e\u0301viter des distorsions de la modulation [16]. La figure 4 illustre l\u2019application de cette technique de calcul des commutations. Il faut cependant ne pas de\u0301passer la valeur maximale du vecteur \u2206\u03a6s, qui est donne\u0301e par : \u2206\u03a6smax = Vsmax \u00b7 Te. (12) Pour un montage e\u0301toile de l\u2019onduleur, on a Vsmax = E/ \u221a 2, ou\u0300 E est la valeur de la tension continue d\u2019entre\u0301e de l\u2019onduleur. La pe\u0301riode de calcul Te fixe l\u2019accroissement du flux statorique. Les valeurs faibles de Te limitent a\u0300 chaque pas de calcul l\u2019excursion du flux, mais cela n\u2019alte\u0300re en rien la dynamique du couple" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001474_bf00533283-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001474_bf00533283-Figure3-1.png", + "caption": "Fig. 3. Projection in the X - - Z plane", + "texts": [ + " Wi th this assumption, the angular moments , m x and my, per uni t length due to the gyroscopic forces are equal to the rate of change of angular momen tum, thus m x = -- m r ~ ~) -- 2 m r 2 ~Q ~'v, (7) my = - - m # ~ + 2 m r 2 ~Q ~, (2) where dots denote differentiat ion with respect to t ime t. Beside the moments due to ro ta t ional acceleration, the shaft is also subjected to inert ia forces due to t ransla t ional acceleration. These result in the following forces per unit length: 82 q~ - - - , , ~ {x + ~ cos ( t? t + 7) } , (3) 02 q y = - - m ~ { y + e s i n ( Q t + y ) } , (4) where e = e(z) is the distance between the geometrical centre and the centre of mass, and Y = Y(z) is the angle fixing the unbalance. 3.2. Conditions of equil ibrium In Fig. 3 the positive directions are defined for forces T x and couples M x act ing on a small element of the shaft in the XZ-plane. We use analogous definitions in the YZ-plane. Apply ing the condit ions of equil ibrium of moments , we ob ta in : OM~ az + m ~ - / 3 , ( 6 + Q w) - - c% /~ = - - Tx, (5) aM~ a--7- + my - - fl, (w - - f2 v) - - c~, w = - - T y , (6) Ing.Arch. Bd. 42, H. 1 (1972) 1% T. Pedersen: On Forward and Backward Precession of Rotors 29 where a~ - - ~(z) and/3, = fi,(z) are damping constants corresponding to an external viscous damping moment , and a rotating internal viscous damping moment , respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001664_j.cma.2010.08.005-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001664_j.cma.2010.08.005-Figure2-1.png", + "caption": "Fig. 2. Uniaxial tension test. (a) Axial directio", + "texts": [ + " In the following, GA stopping criteria options were modified to encounter only for the tolerance on the objective function. GA stops iterating and converges once it finds a best individual scoring less than 10 10. In our case, the mean score between individuals of the same population is noted to be less than 10 5. The same options are used for the identification procedure in the two following examples. For both examples that shall be treated in the upcoming paragraphs, we adopt the uniaxial tension tests shown in Fig. 2. Uniaxial tests are planar tests. By so, when the load is applied only in the e1\u2013e2 plane (Fig. 2) then stress in e3 direction, r3, is null. To build the measured stress tensor Sm introduced in Section 3.1, we use the experimental data extracted from [2] for the HGO model and from [13] for the Fung\u2019s model. HGO constitutive law models the symmetry of the structure by aligning the maximum fibre stretch direction with the mean fibre orientation [15]. The analytic 3-dimensional expression of Cauchy stress derived from the strain-energy function formed by Eq. (8) and (9) is given in [19]. The volumetric preserving parameter was substituted in r1 and r2 after its value was determined by solving stress equation in e3 direction, r3 = 0", + " r1 \u00bc C1 2 k2 1 k 2 1 k 2 2 \u00fe k1 4k2 1c2\u00f0I4 1\u00de exp\u00f0k2\u00f0I4 1\u00de2\u00de h i ; \u00f026:a\u00de r2 \u00bc C1 2 k2 2 k 2 1 k 2 2 \u00fe k1 4k2 2s2\u00f0I4 1\u00de exp\u00f0k2\u00f0I4 1\u00de2\u00de h i ; \u00f026:b\u00de where c and s respectively denote the cosine and the sine of the phenomenological angle b representing the angle between the collagen fibres and the circumferential direction. According to [2], b is set to 43.39 . Figs. 3 and 4 are the response of the HGO constitutive model for the Media layer in the case of the uniaxial tension test. If the sample is loaded in the circumferential direction Fig. 2(b), d1 is a known loading parameter. Hence, by using (26.b) and the material parameters C1, k1 and k2 presented in Table 2, d2 is calculated from the free traction loading condition r2 = 0. Material parameters values given in Table 2 are obtained in [2] thanks to a deterministic inverse procedure. The corresponding strain curve, d2 vs d1 (Fig. 4(b)), shows the toe zone in which collagen fibres start to store strain-energy. The stress component r1 calculated from (26.a) is plotted against d1 in Fig", + " This strain-energy function allows to model arterial layers\u2019 behaviour. In [13], Holzapfel proposed a PI procedure based on deterministic methods to identify the material parameters corresponding to the Fung model. He also presented experimental stress\u2013strain curves for uniaxial tests in terms of 2nd Piola\u2013Kirchhoff stresses vs Lagrangian strains for each of the three constitutive arterial layers \u2013 Adventitia, Media, and Intima (Fig. 6). The corresponding in-plane stresses with respect to the directions shown in Fig. 2 are: S1 \u00bc lf1 \u00bd\u00f02E1 \u00fe 1\u00de2\u00f02E2 \u00fe 1\u00de 1g \u00fe C\u00bdexp\u00f0Q\u00de\u00f02c1E1 \u00fe c3E2\u00de ; \u00f028:a\u00de S2 \u00bc lf1 \u00bd\u00f02E2 \u00fe 1\u00de2\u00f02E1 \u00fe 1\u00de 1g \u00fe C\u00bdexp\u00f0Q\u00de\u00f02c2E2 \u00fe c3E1\u00de ; \u00f028:b\u00de where Q \u00bc c1E2 1 \u00fe c2E2 2 \u00fe c3E1E2. We follow the same procedure as in the previous section to form the so called \u2018\u2018reference data\u201d. The parameters to identify, i.e. l, C, c1, c2 and c3, are substituted into (28.a) and (28.b) by using their given values calculated in [14]. Additionally, in order to build the measured stress vector Sm, the values of the strain couples (E1,E2) are extracted from the strain curves provided in [14] (see also Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002876_1475921713475469-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002876_1475921713475469-Figure5-1.png", + "caption": "Figure 5. Golden section search.", + "texts": [ + " Filtering operation is done by converting the time domain signals into frequency domain and performing a multiplication operation, as shown in Figure 4 W a, b\u00f0 \u00de= F 1 X f\u00f0 \u00dec f\u00f0 \u00de\u00f0 \u00de \u00f04\u00de where X (f ) and c(f ) are the Fourier transforms of time signals x(t) and c(t), respectively, and F 1 is the inverse Fourier transform. Golden section search. An accelerated one-dimensional search is performed in coarse tuning step, as shown in Figure 2. The method we use is the golden section search, which deals with a unimodal objective by rapidly narrowing an interval guaranteed to contain optimum.11 Figure 5 illustrates the idea of the golden section search, where four carefully spaced points are iteratively considered. Leftmost x(lo) is a lower bound on the optimal x , and x(hi) is an upper bound. The function optimum lies between the intervals \u00bdx(lo), x(hi) . Points x(1) and x(2) are the intermediate points. Each iteration determines whether the objective is better at x(1) or x(2), if x(1) proves better, the move direction for the next iteration is left and x(2) becomes x(hi), and if x(2) proves better, the more direction for the next iteration is right and x(1) becomes x(lo)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001709_j.msea.2009.12.008-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001709_j.msea.2009.12.008-Figure10-1.png", + "caption": "Fig. 10. Step patterns in fatigue crack propagation: (a) schematic representation of a a o", + "texts": [ + " 8b)) while the ngle between the coarse lamellae and the crack propagation irection is small (less than 45\u25e6). Otherwise, the crack prefers to ropagate shearing vertical to the colonies. Two types of steps in fatigue crack propagation are observed ig. 9: Steps parallel (Fig. 9a) and perpendicular (Fig. 9b) to the rack propagation direction. These steps occur in a local area of one r two colonies and have short step lengths of no more than 20 m ue to the small size of the lamellar colony. Schematics of these tep patterns in fatigue crack propagation are displayed in Fig. 10. It is reported that the crack growth directions were either parllel to (a1) or perpendicular to (a2) the lamellae orientation in i\u20136Al\u20132Sn\u20134Zr\u20132Mo\u20130.1Si alloy because of the easy slip transission in almost parallel slip direction between b2 on (1 1 0) plane nd a2 on the (0 0 0 1) basal plane [15]. The fatigue crack growth in i\u20136Al\u20132Zr\u2013Mo\u2013V alloy proceeds along the planes either perpenicular or parallel to the long axis of lamellae. The explanation or this behavior is related to prismatic slip along the a1[2 1\u0304 1\u0304 0] nd a2[1 1 2\u0304 0]/a3[1\u0304 2 1\u0304 0] directions [4]. In the present study, the rack grows in the same way considering the prismatic slip affecion and the steps occur (Fig. 10a). However, colony size controls he effective slip length. Small colony size in laser deposited i\u20136Al\u20132Zr\u2013Mo\u2013V alloy results in short slip length and the fine teps in the crack propagation path. The boundaries of the colonies with different orientations mpede crack propagation by fracture of favorite slip plane. When he crack growth encounters the colony boundaries, the relative [ [ [ ngineering A 527 (2010) 1933\u20131937 1937 tilt of fracture planes in neighborhood colony with different orientation will cause the interaction of twist boundaries, forcing the crack propagates along many parallel planes connected by steps (Fig. 10b). Deformation often occurs when these two types of steps are formed. The steps and accompanied tearing pattern in the fatigue crack propagation help to increase the sinuosity and the length of the crack growth path, and hence can lead to more energy consumption during crack growth. Therefore, it can improve the ductility of materials as well as the fatigue crack growth resistance. 4. Conclusions 1. Fatigue crack nucleation in laser deposited Ti\u20136Al\u20132Zr\u2013Mo\u2013V alloy is strongly sensitive to the fine and lamellae" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001653_j.diamond.2010.03.014-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001653_j.diamond.2010.03.014-Figure3-1.png", + "caption": "Fig. 3. Optical image of the finished BDD microelectrode array on sapphire.", + "texts": [ + " Inc.). The microelectrode array was realized by means of the following microfabrication routine: the diamond BDD electrodes have been patterned with a gap distance of \u223c5 \u00b5m by optical lithography and dry etching in an O2/Ar plasma down to the sapphire substrate. Au/Ti ohmic contacts have then been realized. Finally, the surface was passivated with SU-8 except for a circular area of around 16 \u03bcm in diameter exposed to the liquid and the cell, representing the electrochemically active quadrupole. Fig. 3 depicts the finished quadrupole structure viewed by opticalmicroscopy. The device has been packaged as shown in Fig. 4, connecting the ohmic contacts to the measurement board with silver paste and gluing a glass ring to produce the 0.5 \u00b5l chamber for the electrolyte. The amperometric measurements were carried out using an Ag/AgCl counter electrode. After the growth of the electrode materials stack, for removal of residual graphitic phases on the surface, an anodic treatment in the oxygen evolution range was applied in 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001667_jjap.49.04dl04-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001667_jjap.49.04dl04-Figure6-1.png", + "caption": "Fig. 6. Photograph of the experimental equipment.", + "texts": [ + " Each pixel had a selection transistor, which turns on the select signal. The select signal for each pixel was supplied by the integrated shift resistor. The pH and optical signals were output serially after being converted into electrical voltage signals in each pixel by source follower circuits. The output signals from the sensors were processed through an external AD converter before being input to a computer. Computer software converted the signals into a real-time moving image at 5 frames/s. A photograph of the measurement system is shown in Fig. 6. 3.2 Basic characteristics Measurements of the proposed image sensor characteristics were then carried out. To evaluate the device performance, the pH signal was obtained using three standard buffered solutions and the optical signal was applied using lightemitting diodes (LEDs). The reference electrode voltage was maintained high during these measurements. The basic characteristics of the pH and optical sensing are shown in Fig. 7. As is clear from Fig. 7(a), the sensitivity of the output signal obtained was 32mV/pH" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001095_s11044-010-9236-5-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001095_s11044-010-9236-5-Figure15-1.png", + "caption": "Fig. 15 Driving torque applied to crank shaft", + "texts": [ + "2 Four-bar mechanism with large deformable coupler A four-bar mechanism, which consists of two rigid bodies modeled for a crank shaft and a follower and one very flexible body for a coupler, is considered in the second example as shown in Fig. 14. The mechanism is driven by a moment applied to the crankshaft. The Fig. 13 Shear strain at the constrain definition point Fig. 14 Tip displacements driving moment as a function of time is given as [13] M(t) = { 10.00 sin(3\u03c0t), 0 \u2264 t \u2264 0.2778, 465.88e\u221216.32t , t > 0.2778 (53) and the preceding driving function is shown in Fig. 15. The dimensions and the material properties of links used in this example are given in Table 2 [5]. The motion of the fourbar mechanism that exhibits large deformation of the flexible coupler is shown in Fig. 16. The global positions of the midpoint on the flexible coupler obtained using 10, 20 and 30 elements are shown in Fig. 17. It is observed from this figure that the use of 20 elements leads to the convergent solution. In this investigation, a systematic procedure that can be effectively used for modeling joint constraints for the absolute nodal coordinate formulation is developed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000339_j.snb.2008.01.055-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000339_j.snb.2008.01.055-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of the electrochemical re", + "texts": [ + " Besides these applications of PQQ, in this study, QQ was used as an assimilated substrate for TYR electrode to easure pesticide without adding the substrate to enzyme. In ddition, gold nanoparticles (AuNPs) were employed to detect ow-concentration pesticides. For the preparation of substrateound TYR electrode, TYR was immobilized on the PQQ nchored AuNPs by covalent bond. The tethered PQQ was educed by obtaining two-electrons from the electrode and it ould be oxidized to its native form by enzymatic reaction of YR (Fig. 1). The activity of TYR can be maintained by proiding the reduced PQQ, which was generated by reduction at lectrode under a specific potential applied. Eventually the elecrode allows in situ and on site detection of pesticides without ubstrate solution. . Experimental The cell was hooked up to a potentiostat system (Autolab, GSTAT30/GPES, The Netherlands). A polished glassy carbon GC) electrode was used as working electrode. A BAS MF 2030 g/AgCl reference electrode and a Pt wire counter electrode ere also employed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001813_icarcv.2010.5707890-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001813_icarcv.2010.5707890-Figure1-1.png", + "caption": "Fig. 1: Free Body Diagram of a symmetrical quadrotor", + "texts": [ + "ndex Terms\u2014Backstepping, Multi-Loop, Lyapunov stability, Quadrotor I. INTRODUCTION The quadrotor platform is an under-actuated platform, using four rotors as (force) inputs to control the six (coordinate) output states (see figure 1). Research into the quadrotor control problem has led to many potential solutions, several of which have been implemented successfully on hardware testbeds. PID control is used by, for example [1], [2]. For the same problem, a feedback linearization approach was used by [3] and [4]. In addition, sliding mode controllers [5] and \ud835\udc3b\u221e controllers [6], [7] have been used to stabilize and control quadrotor platforms. Finally, another popular nonlinear controller technique, backstepping procedure has been used by [3], [5], [8] and forms the motivation for our present approach", + " The latter system is needed to provide information regarding the attitude and position of the craft with respect to the Earth (observer). The three angle ZYX rotation sequence [11] and the corresponding rotation matrix is given below: R\ud835\udc38\u2192\ud835\udc4f = R\u03a6R\u0398R\u03a8 (1) This gives the total transformation matrix R\ud835\udc38\u2192\ud835\udc4f: R\ud835\udc38\u2192\ud835\udc4f = \u23a1 \u23a3 \ud835\udc9e\u03a8 \ud835\udc9e\u0398 \ud835\udcae\u03a8 \ud835\udc9e\u0398 \u2212 \ud835\udcae\u0398 \ud835\udc9e\u03a8 \ud835\udcae\u0398 \ud835\udcae\u03a6 \u2212 \ud835\udcae\u03a8 \ud835\udc9e\u03a6 \ud835\udcae\u03a8 \ud835\udcae\u0398 \ud835\udcae\u03a6 + \ud835\udc9e\u03a8 \ud835\udc9e\u03a6 \ud835\udcae\u03a6 \ud835\udc9e\u0398 \ud835\udc9e\u03a8 \ud835\udcae\u0398 \ud835\udc9e\u03a6 + \ud835\udcae\u03a8 \ud835\udcae\u03a6 \ud835\udcae\u03a8 \ud835\udcae\u0398 \ud835\udc9e\u03a6 \u2212 \ud835\udc9e\u03a8 \ud835\udcae\u03a6 \ud835\udc9e\u03a6 \ud835\udc9e\u0398 \u23a4 \u23a6 where \ud835\udc9e\ud835\udefc, \ud835\udcae\ud835\udefc is shorthand notation for cos\ud835\udefc and sin\ud835\udefc respectively. Starting with Newton\u2019s second law and using figure 1, the equations of motion can be derived [3]. \ud835\udc40 ( \u02d9\u20d7 \ud835\udc49\ud835\udc4f + \u03a9\u20d7\ud835\udc4f \u00d7 ?\u20d7?\ud835\udc4f ) = \u03a3\ud835\udc39 (2) I\ud835\udc4f \u02d9\u20d7 \u03a9\ud835\udc4f + \u03a9\u20d7\ud835\udc4f \u00d7 I\ud835\udc4f\u03a9\u20d7\ud835\udc4f = \u03a3?\u20d7? (3) \u02d9\u20d7 \ud835\udf01 = J\u03a9\u20d7\ud835\udc4f (4) Here, \ud835\udc49\ud835\udc4f is the velocity vector, \u03a9\u20d7\ud835\udc4f is the body angular velocity vector [\ud835\udc43, \ud835\udc44, \ud835\udc45]\ud835\udc47 , \ud835\udc40 is the quadrotor mass, and 978-1-4244-7815-6/10/$26.00 c\u20dd2010 IEEE ICARCV2010 # \ud835\udc3c\ud835\udc4f the inertia matrix. All terms are expressed in the \ud835\udc4f-frame. Equation 4 converts the body angular vector to a rate of change vector of the Euler angles: [\u03a6\u0307, \u0398\u0307, \u03a8\u0307]\ud835\udc47 . These Euler angles are needed to relate the rotation of the quadrotor to an orientation in space [11]", + " = \u2212\ud835\udc54 cos\u0398 cos\u03a6 +\ud835\udc44\ud835\udc48 \u2212 \ud835\udc43\ud835\udc49 + 1 \ud835\udc40 (\ud835\udc391 + \ud835\udc392 + \ud835\udc393 + \ud835\udc394) (10i) ?\u0307? = [cos\u03a8 cos\u0398]\ud835\udc48 + [cos\u03a8 sin\u0398 sin\u03a6\u2212 sin\u03a8 cos\u03a6]\ud835\udc49 + (10j) +[cos\u03a8 sin\u0398 cos\u03a6 + sin\u03a8 sin\u03a6]\ud835\udc4a ?\u0307? = [sin\u03a8 cos\u0398]\ud835\udc48 + [sin\u03a8 sin\u0398 sin\u03a6 + cos\u03a8 cos\u03a6]\ud835\udc49 + (10k) +[sin\u03a8 sin\u0398 cos\u03a6\u2212 cos\u03a8 sin\u03a6]\ud835\udc4a ?\u0307? = [\u2212 sin\u0398]\ud835\udc48 + [cos\u0398 sin\u03a6]\ud835\udc49 + [cos\u0398 cos\u03a6]\ud835\udc4a (10l) In equation 10c, \ud835\udc36 is a force-to-moment coefficient, converting the (individual) rotor forces into rotor torques. This torque is needed to control the yaw motion. The rotors are setup in two counter-rotating pairs (see figure 1) so that at equal rotor speeds the torques cancel each other. By changing the rotor speeds, a net torque can be generated to stabilize and control the yaw motion. A very important mode of flight for the quadrotor is the hover mode. Without a stable hover mode, it is not possible to hold a certain position. In the hover mode, all speeds, angular rates, and angles are zero and the position states (\ud835\udc4b, \ud835\udc4c, \ud835\udc4d) are constant. The linearized equations (using lower case variables to represent deviations from the equilibrium hover mode) are as follows: " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001590_bf02327752-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001590_bf02327752-Figure6-1.png", + "caption": "Fig. 6~ldeal slip band in a thin uniaxial strip", + "texts": [], + "surrounding_texts": [ + "verge considerably even though their t rends are the same. These observations mere ly ampl i fy the nonl inear nature of the relat ionship be tween thickness change and /o r strain and birefr ingence for polycarbonate as is also indicated in Fig. 1.\nOther differences be tween thickness change and birefr ingence are noted in Fig. 3 even when the var i - ables are small. An obvious explanat ion results f rom consideration of ini t ial local var ia t ion in the th ickness of the mater ia l while in a s t ress-free state. Init ial local variat ions of approximate ly 5 X 10 -4 in. were cer ta in ly present and would not necessari ly be represented by subsequent isochromatics produced by gross yielding. Thus, such local var iat ions could easily produce significant differences in the results shown in Fig. 3.\nHolographic Verification While the qual i ta t ive relationship be tween plastic thickness var ia t ion and plastic isochromatics is clearly established as outlined in the foregoing comments, quant i ta t ive correlat ion was not possible. For this reason, it was felt necessary to have additional qual i ta t ive data to re inforce the mechanical th ickness measurements . In recent years, holography has been shown to be an effective method of measur ing and visual ly observing thickness changes 1244. However, in the present case, a hologram of residual deformations of a permanent ly deformed specimen was desircd, i.e., the specimen shown in Fig. 2. Also, the thickness changes involved represent larger var i - ations than can be observed or recorded in conventional holography. For this reason, the same specimen as previously discussed and shown in Fig. 2 was immersed in an oil bath of near ly the same ref rac t ive index. A double-exposure hologram of the oil bath, with and wi thout the specimen, was taken. This procedure lowers the sensit ivity of the process such that re la t ively large thickness changes can be observed TM.\nThe hologram obtained in the above manner is shown in Fig. 4. Comparison of Fig. 4 and Fig. 2 indicates again that isochromatics in a pe rmanent ly yielded region and holographic in terference fr inges of permanent thickness changes in the same region\nare v i r tua l ly identical. The reader is advised to note that the holographic in ter ference fr inges shown in Fig. 4 were taken wi thout c i rcular ly polarized l ight and, as wil l be subsequent ly demonstrated, the fr inges do not represent isochromatic fr inges which can be obtained using holography 12-t4.\nA second specimen of polycarbonate was machined to be wedge shaped. Care was taken to insure that no residual i sochromat ic- f r inge pa t te rn was induced by the machining process. A double-exposure hologram of the oil tank with and wi thout this stress-free, isochromatic-free , machined wedge is shown in Fig. 5. Obviously, the in ter ference fr inges displayed in Fig. 5 represent thickness change and can be used as a cal ibrat ion of the holographic technique used. Using the cal ibrat ion factor so obtained, thickness changes occurring in the specimen could be calculated f rom the hologram of Fig. 4. This was accomplished at one point and the result is plot ted as point A in Fig. 3 and closely matches the previous ly recorded mechanical measurements . There is, of course, the possibility that the fringes observed in the hologram in Fig. 4 are re la ted to the var ia t ion of re f rac t ive index with the state of stress which produces the thickness change. However such higher order effects are not thought to be large as Fig. 5 would tend to verify.\nAnalytical Observations In the foregoing, it is apparent that a close qual i tat ive correlat ion be tween plastic thickness changes and plastic isochromatics occur in a uniaxial specimen of polycarbonate when stressed beyond the point of tensile instability. However , no reference has been made to the mechanism by which such behavior occurs nor has there been an a t tempt to analyt ical ly interpre t these observations.\nAn analyt ical mode] has been presented by Nfidai 2 to explain s l ip-band formation in metals, i.e., Lfiders' lines. Het6nyi l used the same mode] to explain simi-", + "l a r s l i p - b a n d f o r m a t i o n s in a n y l o n copo lymer . This m o d e l can also be used in t he p r e sen t case.\nF i g u r e 6 r e p r e s e n t s a s chemat i c d i a g r a m of the un i - ax ia l spec imen and its p r e d o m i n a t e slip band which was p r e v i o u s l y shown in Fig. 2. Us ing e l e m e n t a r y elast ic ana lys i s i t is eas i ly s h o w n that ,\n1 ~'x = - ~ (~'x - ~ ' y )\nf r o\n= [1 - - cos 2a - - ~(1 -4- cos 2a) ] (1) 2E\nObvious ly , ~'x va r ies b e t w e e n the two e x t r e m e s ~r O ~r O t e= = ey = and ' = e = = ex --~ . T h e r e exis ts an E E i n t e r m e d i a t e posi t ion such tha t ~'x = 0 and if P0 i s - 1 son's r a t io is t a k e n up as g = -~-, th is pos i t ion wi l l\noccur on a p l ane w h e r e a ~ 35.3 deg. M e a s u r e m e n t of t he o r i en ta t ion of the p r e d o m i n a n t slip p l ane in Fig. 2 y ie lds an ang l e a ~ 35.0 deg. Thus, slip bands seem to occur on p lanes for w h i c h one c o m p o n e n t of the s t ra in is zero. The e x p l a n a t i o n wh ich has b e e n g iven for th is p h e n o m e n a is tha t y i e ld ing can t a k e p lace on these p lanes w i t h o u t cons t r a in t f r o m the s u r r o u n d i n g m e d i u m 1, 2\nWhi l e t h e above e x p l a n a t i o n m a y tend to exp la in t he m e c h a n i c s of s l i p -band fo rmat ion , it does not a id in t h e exp l ana t i on of t he a c c o m p a n y i n g opt ic effect of the m a r k e d increase of f r i n g e - o r d e r n u m b e r in s l ipband regions. Qui te obvious ly , f rom the ea r l i e r sect ion on e x p e r i m e n t a l resul ts , th ickness change in s l i p - b a n d reg ions seem to h a v e a g rea t dea l of inf luence on the obse rved opt ic effect. In fact , these th ickness changes r e q u i r e t h e ex i s t ence of a t h r e e - d i m e n s i o n a l s tress s ta te in a reas of gross y ie ld ing. Howe~e r , u~ing n o r m a l a r g u m e n t s of photoe las t ic theory, i t ~ r e a s o n a b l e to in fe r tha t the s t ress no rma l to t h e p l ane of the spec imen, ez, does\" not inf luence\nt he optic effect. Also, i t is a p p a r e n t tha t the stress caused by p e r m a n e n t l a t e r a l con t rac t ion is n e a r l y zero, ~x ~'~ 0. As a resul t , t he s tress s ta te in a slip band can be t h o u g h t of as u n i a x i a l e v e n beyond the y ie ld poin t of t h e mate r ia l . The re fo re , in t h e fo l lowing, a un i ax i a l - s t r e s s s ta te w i l l be assumed.\nTo aid f u r t h e r discussion, idea l i zed s t r e s s - s t r a i n - optic d i ag rams a r e shown in F i g . 7. A s t ress -op t ic l aw w h i c h w o u l d be consis tent w i t h photoe las t ic i ty and stil l b e ' v a l i d b e y o n d the e las t ic l imi t w o u l d be,\n: f% y + %.p. 1 -- (2) fe c /\nw h e r e ~ is the ax i a l s t ress at a point , ~y.p. is the y ie ld stress, nc is the f r i nge order , d t he th ickness at the po in t in ques t ion , ]% the elast ic m a t e r i a l - f r i n g e value , and ]% the p las t ic m a t e r i a l - f r i n g e va lue . Equa t i on (2) is ana logous to a s t r a i n - h a r d e n i n g s t ress -s t ra in l aw of the type,\ne = E v ~ + % . v . 1 Ee (3)\nw h e r e 9 is the s t r a in a t a point , E e is an elast ic m o d u - lus and Ev is a p las t ic modu lus . The basic ques t ion\n470 I October 1971", + "to ask here is whether the idealized s t ress-s t ra inoptic diagrams shown in Fig. 7 are realist ic representations of the behavior of polycarbonate. F rom consideration of Fig. 1, it is possible to infer that idealized stress-strain response is a close approximat ion to the behavior of polycarbonate. The only question is the magni tude of Ep. Here it is assumed that E~ ~ 0 but is small compared to E e, i.e., Ep < < E e. In other words, while strain hardening is assumed to take place, it is also assumed to be quite small. With these assumptions and with the aid of Fig. 1, it is possible to construct the idealized stress-optic response shown in Fig. 7 which would be quite similar to the idealized stress-strain response shown there also. Equat ion (2) is a logical consequence of this idealized response and can be rewr i t t en as,\n( ) ne = . . . . 1 ( ~ - - ~y.p.) (4)\n--d- 5\"e fec\nThe interpreta t ion of the stress-optic effect represented by eq (4) depends on whe the r the initial thickness, di, or the current thickness, dc, at a mate - r ial point is used. The proper choice is to use the current thickness as normal photoelastic theory dictates, even though in photoelast ici ty the init ial thickness is most often used because there is l i t t le discrepancy be tween dl and dc in the elastic range. Use of both thicknesses are shown in Fig. 7 and it can be observed that a small decrease in thickness represents a re la -\nt ively large increase in for the same stress\nlevel. In fact, it is in tui t ively obvious that, should the thickness throughout the throat of a tensile specimen remain uniform, then the fr inge order would also be uniform. If the thickness varies over a slipband region, a mul t ip le - f r inge pat tern would be expected and fringes would then tend to be associated with lines of constant thickness. The earl ier exper i - menta l evidence of Fig. 2 and Fig. 3 reinforces these remarks.\nThe residual fr inge pat tern of Fig. 2 after unloading can be predicted using eq (4) by assuming elastic unloading resul t ing in the equation,\n(5) /\nAt this point, it is also worth noting that uniaxia l - tension holographic interference fringes of thickness change for elastic deformation can be represented by the equation 12,\n= Iep (6)\nwhen n~, is the fringe order, d is the current thickness, and )~r, is the isopathic mate r ia l - f r inge value. While it is not known whe the r eq (6) can be extended to include inelastic deformations by rewr i t ing in the form of eq (2), it is clear that holographic fr inges and isochromatic fringes would have the same character in uniaxial tension. Therefore, it is not surprising that the hologram of Fig. 4 is similar to the isochromatic pa t te rn of Fig. 2.\nDiscussion The exper imenta l results of the present invest iga-\ntion indicates good qual i ta t ive correlat ion be tween lines of constant- thickness change and isochromatic fringes in a s l ip-band region of a uniaxia l tensile specimen. The analyt ical observat ions presented also indicate that such a relat ionship should exist. The formula ted stress-optic law has only been used as a means of explaining and discussing the relat ionship be tween thickness and isochromatic measurements . A strain-opt ic law could be used for the same pur - pose, par t icular ly since thickness changes are re la ted to the state of strain in a one- or two-d imens iona l stress state. However , due to the s imilar i ty be tween the stress-optic representa t ion used here in and more rigorous elasto-plast ic s t ress-strain equations, it would seem conceptual ly possible to fo rmula te a theory of photoplast ici ty for three dimensions which would closely paral le l the extensive theory of plasticity.\nThe holographic results may have possible fu ture application in elastic and /or plastic stress analysis. As suggested by Hovanesian, et a112, holographic int e r fe romet ry could be used in conjunct ion with twoor three-d imens ional stress freezing, even for the re la t ive ly large deformat ions involved. This could be accomplished by using an oil bath of near ly the same refract ive index to lower the sensi t ivi ty of the process.\nA c k n o w l e d g m e n t\nThe author grateful ly acknowledges the financial and laboratory support supplied under N A S A contract no. NGR 47-004-051. Apprecia t ion is also extended to Dell Williams, C. A. Hermach, and Richard Brown of NASA-Ames , Moffett Field, California, for their many helpful discussions and support. The wr i te r is also indebted to Dr. Brown for use of his holographic equipment and to my colleagues at VPI for their helpful criticisms and suggestions.\n1. Hetdnyi, M., \"'A Study in Phetoplasticity,\" Proc. First Natl. C(mtL, Appl. Mech. (1952). 2. Nddal, A., Theory of Flow and Fractnre of Solids, Vol. 1, McGraw-Hill, N e w York (1950). 3. Frocht, M. M., ed., Symposium on Photoelasticity, The MacMillan Company (1963). 4. Hunter, A. R., \"'Development of a Photoelasto-plastic Method to Study Stress Distributions in the Vicinity of a Simulated CrackPhase I and I I ,\" Lockheed Missiles and Space Company Report 4-05-65-11 (Oct. 1965).\n5. Frocht, M. M. and Thomson, R. A., \"'Studies in Photoplasticity,\" Proc. 3rd U. S. Natl. Cong., Appl. Mech. (1958).\n6. Bayoumi, S. E. A. and Frankl, E. K., \"'Fundamental and Relations in Photoplasticity,'\" Brit. ]nl. of Appl. Phys., 4 (1953).\n7. Fried and Shoup, N. H , \"'A Study in Photopiasr The Photoelastic Effect in the Region of Large Deformation in Polyethylene,\" Proc. 2nd U. S. Natl. Cong., Appl. Mech. (1954).\n8. Brill, W . A., \"'Basic Studies in Photoplasr PhD Thesis, Stan]ord University (1965).\n9. Gurtman, G. A., Jenkins, W . C. and Tung, T. K., \"'Characterizat{on of a Bi'rc[ringent Material for Use in PhotoeIasto-Piastieity,'\" Douglas Report SM-47796 (January 1965).\n10. Whitfield, J. K., \"'Characterization of Polyearbonate as a Photoelasto-Plastia Material,\" PhD Thesis, VPI (Dec. 1969).\n11. Brinson, H. F., \"'The Ductile Fracture of Polycarbonate,'\" EXPEnIMENTaL ~ECnANZCS, 10 (2), 72-77 (Feb. 1970).\n12. Hovanesian, 1. D., Brcic, V. and Powell, R. L., \"'A New Experimental Stress-Optic Method: Stress-Holo-Interferometry,'\" EXPEtLIMENTAL MECHANICS, 8 (8), 362-368 (Aug. 1968),\n13. Fourney, M. E. and Mate, K. V., \"'Further Application of Holography to Photoelastieity,'\" EXPERIMENTAL MECHANXCS, 10 (5), 177-186 (May 1970).\n14. Sampson, R. C., \"'Holographic-interIerometry Applications in Experimental Mechanics,\" EXPERIMENTAL MEC:-IANICS, 10 (8), 313- 320 (Aug. 1970).\nE x p e r i m e n t a l M e c h a n i c s I 471" + ] + }, + { + "image_filename": "designv11_3_0002277_icit.2012.6210043-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002277_icit.2012.6210043-Figure6-1.png", + "caption": "Fig. 6. Motor Identification Test-Bench", + "texts": [ + " MODEL EXTRACTION (17) For the 6-DOF model extraction process no mathematical simplifying assumptions have been made. Instead, the MAT LAB environment has been utilized where an M-File based S Function has been programmed, by solving equations (3),(5) and (4),(6) with the following input and output vectors: It must be noted that the aforementioned dynamics concern the rigid-body dynamics alone. In order to derive a detailed system model special attention must be paid in modeling the system's actuation. Therefore, an experimental test-bench illustrated in Figure 6 consisting of a force measuring device and a rotational speed measuring module was utilized, in order to obtain a thrust-to-rpm measurement table as well as a rotational speed step response plot. The rotors' transfer functions were simulated with a I't order transfer function and were found to be: Rotor 1 Rotor 2 Rotor 3 1 hl(s) = 0.082035s+ 1 1 h2(S) = 0.082035s+ 1 1 h3(S) = 0.04093s+ 1 (19) (20) (21) In Figure 7 the step response plot for rotor 1 is depicted. Because we aim to extract a linear subsystem model, all measurements were taken around the rotor operation region that could achieve stable hover, namely where: 3 [Tf = GB " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000996_j.jsv.2009.11.014-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000996_j.jsv.2009.11.014-Figure3-1.png", + "caption": "Fig. 3. A curved beam with the uniform equilateral triangle cross section.", + "texts": [ + " (13) reduces to m \u20acux \u00bcQx 0 ksQZ\u00fekZQs\u00fepx; m \u20acuZ \u00bcQZ 0 kxQs\u00feksQx\u00fepZ; m \u20acus \u00bcQs 0 kZQx\u00fekxQZ\u00feps; rIx \u20acjx \u00bcMx 0 ksMZ\u00fekZMs QZ\u00femx; rIZ \u20acjZ \u00bcMZ 0 kxMs\u00feksMx\u00feQx\u00femZ; rG \u20acjs 00 \u00ferIp \u20acjs \u00bcMs 0 kZMx\u00fekxMZ\u00fems; (14) in which G\u00bc Z Z w2 dxdZ; (15) where m\u00bc rA is the mass per unit length of the beam, A, Ix, IZ and IP are the cross-sectional area, the second moments of area with respect to the normal axis and to the binormal axis, and the torsional moment of inertia of the cross-section, respectively. To verify the theoretical formulations in previous section, a curved beam with the uniform equilateral triangle cross section is chosen, as shown in Fig. 3, as a model in computation. For the structure, y, ks and kx in Eq. (3) all take zero, and kZ is 1/R. First, we consider out-of-plane free vibration of the beam. In the case, corresponding three equations, the second, the fourth and the sixth equations in Eq. (13), need to be considered. These equations are uncoupled from the rest of the system, and can be expressed in terms of only the three independent displacement functions u\u0302Z, j\u0302x and j\u0302s below rA \u20acuZ\u00f0s; t\u00de \u00bc GZGAuZ 00 \u00f0s; t\u00de\u00feGZGAjx 0 \u00f0s; t\u00de; rIx \u20acjx\u00f0s; t\u00de \u00bc GZGAuZ 0 \u00f0s; t\u00de\u00fe\u00f0k2 ZGIP\u00feGZGA\u00dejx\u00f0s; t\u00de EIxjx 00 \u00f0s; t\u00de kZ\u00f0EIx\u00feGIP\u00dejs 0 \u00f0s; t\u00de\u00fekZGD1js 0 \u00f0s; t\u00de; rIP \u20acjs\u00f0s; t\u00de rG \u20acjs 00 \u00f0s; t\u00de \u00bc kZ\u00f0EIx\u00feGIP\u00dejx 0 \u00f0s; t\u00de\u00fekZGD1jx 0 \u00f0s; t\u00de EGj0000s \u00f0s; t\u00de\u00feGIPjs 00 \u00f0s; t\u00de 2GD1js 00 \u00f0s; t\u00de Ljs 00 \u00f0s; t\u00de k2 ZEIxjs\u00f0s; t\u00de: (16) in which D1 \u00bc Z Z qw qx Z qw qZ x dxdZ; L\u00bc G\u00f0D1\u00fek2 ZG\u00de: For the harmonic vibration with frequency o, introduce the following dimensionless quantities lB1 \u00bc R2ro2 G ; lD1 \u00bc R2GA EIx ; lD2 \u00bc R2\u00f0k2 ZGIP\u00feGA\u00de EIx ; lD3 \u00bc R2ro2 E ; lD4 \u00bc EIx\u00feGIP EIx ; lD5 \u00bc GD1 EIx ; lF1 \u00bc R2\u00f0EIx\u00feGIP\u00de EG ; lF2 \u00bc R2GD1 EG ; lF3 \u00bc R2GIP EG ; lF4 \u00bc 2R2GD1 EG ; lF5 \u00bc R2L EG ; lF6 \u00bc R2ro2 E ; lF7 \u00bc R2EIx EG ; lF8 \u00bc R4ImPo2 EG ; (17) then Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003602_tmag.1970.1066758-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003602_tmag.1970.1066758-Figure4-1.png", + "caption": "Fig. 4. Measured field distribution transverse to direction of motion.", + "texts": [ + " 3 shows the field profiles, which were measured in azimuthal direction and normed to I, for the 2-pole and the 18-pole magnet system. In the case of the 2-pole system, the field strength exhibits pronounced peaks above the poles while in the case of the 18-pole system i t has to all intents and purposes a straightforward sinusoidal shape. The remaining systems have field profiles ranging between these two extremes. Table I gives the numerical factors for different 6 values in the case of the 2-pole and the 18-pole systems. Fig. 4 shows the field strength measured in the %pole system in radial direction. A s the number of poles decreases, the field distribution becomes m.ore symmetrical. With the increase of the number of poles, the field strength maximum shifts towards the right in the direction of the disk center. Furthermore, Fig. 4 contains three functions which are used in the calculation discussed in Section 111, namely, a rectangular function, a cosine, and a cosine-squared curve. It can be seen that the two cosine functions approximate the measured field distribution very well. 111. CALCULATION The formulas for the braking force were derived to relate to a linear magnet system [SI. Fig. 5 shows the buildup of the system which has infinite extension in the x direction. A single plate of thickness d, width b and conductivity U , is interposed between two rows of permanent magnets moving relative to the plate a t velocity v", + " 2, JUNE 1970 has not been taken into account. It can be seen that the braking forces measured directly are well described by one of the two force curves which were calculated with the aid of the cosine function according to (6) or ( 7 ) . The force curve obtained for the rectangular function according to (5) is, however, substantially above the measuring points. This was to be expected as in this case the field strength at the site of the eddy-current disk exceeds the measured field strength by a substantial amount (Fig. 4). Figs. 6 and 7 further contain force curves which were calculated with the aid of the measured field-strength values according to the rule-of-thumb formula [ 2 ] F = kapo2vclnafiH2 (8) where n is the number of poles, a and @ are the dimensions of the magnet pole, and H i s the mean field strength above 263 the magnet poles which in the present case amounts to about 88 percent of the maximum field strength H,,, (Table I). k < 1 is a corrective factor and normally taken to be 0.3 in the case of electricity meters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002211_j.mechmachtheory.2012.01.020-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002211_j.mechmachtheory.2012.01.020-Figure4-1.png", + "caption": "Fig. 4. Reenactment of the gear-cutting device (a) while cutting a tooth, (b) while rotating the disk to a new tooth position and (c) the measuring tool while dividing the disk for a new tooth.", + "texts": [ + " Taking into account that the basic data available to the craftsman was the number of teeth of each gear, a possible method that he could have devised is the following: (a) By multiplying the number of teeth of each gear by a constant number (something similar to the module used nowadays) he could calculate the tip diameter of each gear. (b) He would then cut a bronze disk for each gear, whose diameter was equal to the tip diameter. (c) He would place each disk in a specially constructed device and would mark the location of the teeth on the disk, using a measuring tool (see Fig. 4c). This measuring tool had a double-pointed edge (a gap) which corresponded to the chord length (the distance between two successive teeth) of each gear. (d) He was able, thus, to divide the outside diameter of each disk into the specific and correct number of teeth. (e) Using a spring-plate to hold down firmly each disk on a table, he could file each tooth, using a triangular file (see Fig. 4a), up to the required depth (corresponding to the root diameter of the gear). A stop-mechanism (see Fig. 4b) was used to stop the grinding to the required depth. (f) Each time that a tooth was cut, the craftsmanwould release the spring-plate and rotate the disk to the next position shown by the measuring tool. (g) The next tooth was filed and the process was repeated from item (e). 3. Assessment of the geometrical dimensions of the gears assuming the use of a triangular file In order to confirm the above mentioned scheme, it should be proved that the root angles in all gears are equal and that there exist a fixed ratio between the number of teeth and the diameter of each gear (its \u201cmodule\u201d)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000859_tmech.2009.2032180-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000859_tmech.2009.2032180-Figure5-1.png", + "caption": "Fig. 5. Method for checking for collisions of the segment.", + "texts": [], + "surrounding_texts": [ + "The path obtained in the previous section may not be a smooth one; it may include detours and unexpected discontinuity of velocity. Our smoothing algorithm comprises two steps: the shortcut path is found in the first step and spline interpolation is applied in the second step. In both steps, it is essential to avoid unnecessary collisions of links. The shortcut algorithm is a simple extension of the one proposed in [25], where the time parameter is newly considered. Since the method of collision checking of a segment is exactly the same as Algorithm 4, we mainly explain spline interpolation in this section. The path of the robot\u2019s configuration obtained by the shortcut operation may still include the discontinuity of velocity since two milestones are connected by using a straight-line segment. In the case of a humanoid robot, the discontinuity of velocity must be avoided since the robot may fall down due to the effect of a large acceleration. To overcome this problem, we apply b-spline interpolation. An overview of the spline interpolation algorithm is shown in Fig. 6. Fig. 6(a) shows the path obtained by the shortcut operation. Then, as shown in Fig. 6(b), we split the trajectory between q(ti) and q(tj ) into n segments S1 , S2 , . . . , Sn . By applying bspline interpolation, we can obtain a set of curved trajectories. Fig. 6(b) shows the case in which the trajectory is split into n segments. We check for collisions of the curved segment from S1 to Sn . A collision-checking algorithm for the curved segments can be obtained by simply modifying Step 1 of Algorithm 4 where mi and mj are connected by using a curved trajectory instead of a straight line. If the collision occurs in Si , we consider adding an additional node to Si and splitting Si into Si1 and Si2 . Fig. 6(b) shows the case where collision occurs at S2 . Then, Fig. 6(c) shows that an additional node is inserted to S2 and S2 is split into S21 and S22 , After checking for collision of Si , we check for collision of Si+1 . If we finished checking for collisions of Sn , we return to S1 . We iterate this operation until we obtain a collision-free trajectory connecting q(ti) to q(tj ). If the difference between the shortcut trajectory and the splined trajectory reduces as the number of node increases and if the minimum distance between the shortcut trajectory and the obstacle is greater than 0, we can obtain a smooth and collisionfree path by using the spline interpolation algorithm. As far as we tried, as the number of nodes increased, the difference between the shortcut trajectory and the splined trajectory decreased. However, the same may not necessarily be true if we use spline interpolation." + ] + }, + { + "image_filename": "designv11_3_0002113_09544062jmes2181-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002113_09544062jmes2181-Figure7-1.png", + "caption": "Fig. 7 (a) Film thickness; (b) pressure distribution; and (c) lip distortion [23]", + "texts": [ + " The same general approach is taken in reference [23] as in references [20] and [21] (full film lubrication, smooth shaft, Reynolds equation with cavitation, flooded boundary conditions, and iteration to handle coupling between fluid mechanics and deformation), but both the Reynolds equation and the axisymmetric seal deformation (including the asperity deformation) are solved with on-line finite-element analyses. This model predicts film thickness distributions consistent with references [20] and [21], but the predicted reverse pumping rates are higher. Typical predicted film thickness and pressure distributions and asperity distortions are shown in Fig. 7. In an extension to reference [23], the nonNewtonian shear-thinning behaviour of the sealed fluid is considered in reference [24]. A Gecim law equation (a form of a double Newtonian model) is used for the fluid viscosity. The results indicate that when compared with the Newtonian predictions, the film thickness, power loss, and reverse pumping rate are all reduced by small amounts. The models described earlier are all isothermal, with properties usually evaluated at sump or ambient temperature. However, as heat is generated in the sealing zone, it is expected that temperatures will be elevated, especially the temperature of the fluid in the sealing zone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003695_s00170-015-7533-0-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003695_s00170-015-7533-0-Figure5-1.png", + "caption": "Fig. 5 a Component roll and b sintered components", + "texts": [ + " Figure 4 illustrates the merging of the neighbouring filled holes necessary to produce a continuous component in SW isometric view. Therefore, object ribbon height and processing resolution will be decreased at least by four and two times as a result of pressure roller work. The addition of print head with scanning system makes it possible to control and correct ribbon perforation. It is achievable to create object consisting of different powders in one roll simultaneously, if print head, upper and lower bunker, etc., were added for each powder accordingly. When the whole component roll is rewound (Fig. 5a), it is ready for a sintering machine. Depending on object powder properties (plastic, metal, ceramic, grain size and other), conventional heating, micro-wave, induction sintering, etc., are realizable in an atmosphere of various gases and pressures. At the end of the manufacturing process, it is required to remove the ribbon mechanically by air flows or dissolving, e.g. by water (Fig. 5b). To radically increase productivity, it is possible to sinter several component rolls simultaneously. The RPS technology can be used to process almost any material available as a powder, and the powder particles tend to sinter when heat is applied. The mechanical properties of parts printed with usual layerby-layer additive manufacturing technologies vary depending on the printed direction. Parts printed with RPS are much more like sintered parts. RPS produces consistent and predictable mechanical properties, creating parts that are smooth on the outside and solid on the inside" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002373_s0965545x10120138-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002373_s0965545x10120138-Figure1-1.png", + "caption": "Fig. 1. Diagram of the measuring unit of a rheometer: (1) drive of forced torsion oscillations; (2) MCE sample (tablet), and (3) bed.", + "texts": [ + " The second material was filled with a mixture of two iron powders: R 10 carbo nyl iron (a particle size of 3\u20135 \u00b5m) and powdered iron (a particle size of 40\u201350 \u00b5m) at a 65 : 35 weight ratio; the total content of the magnetic filler was 83 wt %. The main characteristics of the samples are listed in Table 1. All main studies of the viscoelastic properties of the MCE samples were performed with a Haake Rheo Stress 150 rheometer (manufactured by the Thermo Scientific) in the dynamic mode of forced torsion oscillations with a controlled load (the torque) varying according to a harmonic law. We used a plane\u2013plane measuring cell; its sketch is shown in Fig. 1. Cylindri cal MCE samples 2 with a diameter of 10 mm and a thickness of 4\u20135 mm were fixed on stationary lower bed 3 to avoid slipping. Upper plane 1, which was pressed against the upper end of the sample and con nected to the rotor, underwent forced torsion oscilla tions. The frequency and amplitude of the external loading were varied. Unless otherwise specified, the applied shearing stress was 100 Pa for a softer sample and 6\u20131000 Pa for a more rigid sample. Under these loads, the maximum strain of the samples was in the range 1\u201310%", + " To obtain objective data, we modified the technique by abandoning the correction with an unknown algorithm provided by the manufacturer. The primary data for us were \u201cdirect\u201d recordings of the mechanical impedance of a sample that were obtained ( )G f with the device with deactivated correction. Although, as noted above, these data contain a large resonant contribution, the rest of the information on the dynamic behavior of the MCEs in them is distorted minimally. By schematizing the measurement process as shown in Fig. 1, we assume that the lower end of the sample is slip free fixed on the device bed with an infi nite weight and that the upper end is slip free adjacent to the rotor heel and exactly repeats its motion. With this setup, periodic shear oscillations are generated in a tablet of the elastomer. Here, the lower surface of the elastomer is stationary, while the amplitude of rota tional motion on the upper surface has a maximum value that coincides with the angular displacement of the rotor, which is assumed to be a perfectly rigid body", + " Because, in the absence of a field, an MCE is not magnetized and is isotropic, all the new contributions must be bilinear (even with respect to H) convolutions of the field components with the strain tensor e. As shown in [33], there are three linearly independent combinations of this type: , , and , where the boldface straight symbols denote tensors and the boldface italic \u22c5 \u22c5eH H \u22c5 2e eH \u22c5 2( )e H 1350 POLYMER SCIENCE Series A Vol. 52 No. 12 2010 STOLBOV et al. quantities stand for vectors. The addition of these con tributions with the numeric coefficients and , respectively, into the model free energies of the MCE elastic components (Fig. 1a) and the subsequent dif ferentiation with respect to the strain tensor determine the modified elastic stress tensors that depend on the applied field. Viscous stress tensor \u03c3visc must likewise obey the above rules of symmetry except that it does not contain a contribution proportional to . Let us denote the coefficients of additives to \u03c3visc through and . As a result, we obtain the extended PT model, which now comprises seven equations: 1 3\u2212\u03b1 1 3\u2212\u03b2 \u22c5 \u22c5eH H 1\u03b3 3\u03b3 (10) For the harmonic torsion oscillations of an MCE tablet magnetized by a constant field perpendicular to its plane (i", + " Neither of them leads to a better agreement between theory and experiment than the simple 2PT approximation. In the analysis of the magnetorheological effect, we introduced into the 2 PT model the assumption of the quadratic dependence of each parameter on the value of . This condition is not strictly mandatory, because, as noted above, only the response to field of the sample as a whole is observable (measurable). At first glance, the hypothesis on the dependence of all model parameters on is redundant. In fact, for the geometry of our experiment (Fig. 1), as is clear from formulas (12), we can determine only the sums of the coefficients: , \u2026 . However, with variation in the directions of tension\u2013compres sion and magnetization of the sample, these combina tions change. For example, during uniaxial stretching, the value of is significant. Therefore, it is possible to unambiguously find the dependence of any of the 2PT model parameters on the applied field or the absence of this dependence only after a detailed experimental study of the MCEs. The low frequency (0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000520_978-0-387-93808-0_33-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000520_978-0-387-93808-0_33-Figure5-1.png", + "caption": "Fig. 5 Architecture of Laser Pointer", + "texts": [ + " \u2022 Second, when multiple users are interacting, with two-steps associating methods described in this subsection, our system can identify different laser pointers and support multi-user collaboration. \u2022 Last but not least, the laser pointer emits its identity through radio frequency during interaction. The system receives it and treats different users separately. A common laser pointer integrated with three additional buttons and wireless communication modules is introduced as the interaction device in our system. In ad- dition, to receive the RF signal emitted by the laser pointer and transmit it to the computer, a new hardware called the Receiver is also introduced. \u2022 The laser pointer Fig. 5 shows the architecture of it. There are totally three additional buttons on the laser pointer, On/Off button, Right button and Left Button. Their functions are described as Table 2. On/Off button: This button is a switch, not only for turning on the laser pointer but also for broadcasting the user\u2019s identity through wireless communication modules. If the button is down, the laser beam is emitted and the ID of the laser pointer is broadcasted through radio frequency at the same time. The system receives the ID through the Receiver" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003054_rob.21462-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003054_rob.21462-Figure11-1.png", + "caption": "Figure 11. Navigation command: relative distance and heading between each waypoint.", + "texts": [ + " Calculating the evaluation metrics for all candidate paths, the most feasible path between them is determined such that the path has the minimum value of the evaluation metric. The most feasible path, which is determined with the approach described in the above sections, is discretized into several waypoints. A navigation command NavCmdi = (di, \u03c8i), consisting of the relative distance di and heading \u03c8i from one waypoint to the next, can be geometrically calculated based on the terrain map data (Figure 11). Then, the rover mobility executes the sequential navigation commands from a start point to a goal point while traveling through each waypoint. An odometry using wheel encoders is a traditional approach to measure the distance traveled. However, it may not be reliable when the rover travels on sandy loose terrain where the rover experiences wheel slippage, resulting in incorrect calculation of the distance traveled with respect to the wheel rotations. The errors then accumulate over time and eventually degrade the accuracy of the position estimation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003620_s11431-011-4544-4-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003620_s11431-011-4544-4-Figure2-1.png", + "caption": "Figure 2 The connection of all coordinate systems.", + "texts": [ + " (1) and start computation on the next loading step. If some elements are slack or wrinkled, the stresses and the constitutive matrix should be modified according to the procedure in section 4. If there are some wrinkled areas in the membrane structure, the constitutive matrix in the local coordinates in eq. (5) and the principal stresses obtained from eqs. (16) and (17) should be modified before the next iteration. Denoting the angle between the local coordinates (x, y) and the principal stress coordinates (1, 2) as (Figure 2) and the angle between the principal coordinates of the material (W, F) and (1, 2) as , we have the following relationship: 2 1 tan ,xy xy x y (20) . (21) According to eq. (5), the constitutive matrix in the (1, 2) plane is formulated as 11 12 13 T 21 22 23 31 32 33 . D D D D D D D D D D T DT (22) Thus, the principal stress vector can be expressed as the product of the constitutive matrix and the strain vector (i.e., T 1 2 12 ,j e e e e where 12 0e due to the tension- shear coupling effect) in the ),( 21 plane 1 11 12 13 1 2 21 22 23 2 31 32 33 12 ", + " The maximum displacement occurs at the center of the roof with a value of 501.5 mm when the loading is up to 13.14 kN/m2. The shape of the structure is changed from the negative Gaussian curved surface to a positive one at this moment. The relationship of load and displacement at this location is shown in Figure 7, which indicates that the displacement increases with an increase in the loading. Moreover, the rate of increase is the largest when the membrane begins to wrinkle at about 0.66 kN/m2 loading. The principal stress history of element 92 (Figure 2) is presented in Figure 8, which shows that the maximum principal stress of the structure increases with loading while the minimum principal stress decreases first then increases with loading. The value of the minimum stress is negative when the loading reaches 0.66 kN/m2. It denotes wrinkling of the corresponding membrane element. When the load is greater than 1.20 kN/m2, the value of the minimum stress begins to increase and it becomes positive when the load gets to 3.00 kN/m2. This suggests that the slack element changes back to taut again" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000426_s021984360800142x-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000426_s021984360800142x-Figure6-1.png", + "caption": "Fig. 6. Visualization of the \u201cconstant velocity\u201d model for smooth motion.", + "texts": [ + " (5) Explicitly, the camera\u2019s state vector xv comprises metric 3D position vector rW , orientation quaternion qWR, velocity vector vW and angular velocity vector \u03c9R (a total of 13 parameters). Feature states yi are 3D position vectors. In the extended Kalman filter (EKF) prediction step, a model for smooth motion anticipates Gaussiandistributed perturbations VW and \u2126R to the camera\u2019s linear and angular velocity at each time step \u2014 modeling motion with a generally smooth character. The explicit process model for motion at a time step \u2206t is fv = rW new qWR new vW new \u03c9R new = rW + (vW + VW )\u2206t qWR \u00d7 q((\u03c9R + \u2126R)\u2206t) vW + VW \u03c9R + \u2126R . (6) Figure 6 illustrates the model\u2019s potential deviations from a constant velocity trajectory. Implementation requires calculation of the Jacobians of this function with respect to both xv and the perturbation vector. 6.2. Taking walking information into account Providing an EKF-based SLAM system with the robot\u2019s motion model is not new, but the specificity of the humanoid robot makes it interesting. Indeed, the oscillation In t. J. H um an . R ob ot . 2 00 8. 05 :2 87 -3 10 . D ow nl oa de d fr om w w w .w or ld sc ie nt if ic " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002252_13552541111113844-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002252_13552541111113844-Figure3-1.png", + "caption": "Figure 3 Schematic diagram of experimental setup", + "texts": [ + " In order to control the deposition layers\u2019 thickness and height of thin-wall metal parts since it impacts the quality of the products and reduce the surface unevenness caused by the fluctuation of process parameters and system error, as shown in Figure 1, the standoff distances H is also an important factor by experimental observations in Figure 2. In the paper, the effect of the relationship between the standoff distance and the powder focus length on the surface unevenness, the layer thickness and height of thin-wall metal parts was studied, and the steady standoff distance can be acquired during the process by theoretical calculation according to the experimental rule. The experiments were carried out with the setup shown in Figure 3. The system includes a Nd: YAG laser with a 1 kW maximum output power, a three-axis CNC machine and a powder supply system. The whole setup is a typical open-loop DMD system. The laser beam was guided to the workstation through an optical fiber and focused by an optic with a 160-mm focal length. The powder used was stainless steel 316L with spherical shape and the diameter of 15-45mm, which was supplied by sub-institute of Metal Materials of Beijing General Research Institute of Mining and Metallurgy" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000277_s12283-008-0002-3-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000277_s12283-008-0002-3-Figure2-1.png", + "caption": "Fig. 2 Eulerian angle convention. The rotational sequence of motion follows the order of coordinate axes Z, y0 and x0 0. Joint angles q1, q2 and q3 were defined as the rotational angles about Z, y0 and x0 0 axes of the right-handed reference frame, respectively", + "texts": [ + " The model involved the upper trunk rotations about three axes and the throwing shoulder, elbow and wrist joints including 7 degrees of freedom (Fig. 1). Each joint angle of the model was calculated using the Eulerian angle convention described by Chao [4]. The Eulerian convention is defined as the rotational sequence of motion that follows the order of coordinate axes Z, y0 and x00, and the Eulerian angles q1, q2 and q3 represent as rotational angles about Z, y0 and x00 of the right-handed reference frame, shown in Fig. 2. For the upper trunk, the counterclockwise (+) and clockwise (-) rotation hY, medial (+) and lateral (-) lean hP, and posterior (+) and anterior (-) lean hR were defined as the Eulerian angles q1, q2 and q3 about Z, y0 and x00, respectively. The neutral position of the upper trunk reference frame was coincident with the global reference frame in three directions. In this case, the unit vectors corresponding to the directions of Z, y0 and x00 were defined as the joint axes of the upper trunk counterclockwise/clockwise rotation (sY), medial/lateral lean (sP) and posterior/anterior lean (sR), respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002668_cdc.2013.6760297-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002668_cdc.2013.6760297-Figure4-1.png", + "caption": "Fig. 4. Airframe nomenclature.", + "texts": [ + " By working with the Standard Form of Figure 3 and exploiting the nature of the performance criteria (5), this algorithm is amenable to high-performance implementations that provide an excellent compromise between flexibility of the design tool and tractability of the underlying optimization program. A suitable solution x of (6) determines the coefficients Kjl for the gain surface (4). As with the conventional approach, the resulting gain-scheduled controller comes with no global guarantees of performance, so validation on a finer \u03c3 grid and nonlinear simulations remain necessary to fully qualify the results. This section applies the \u201dGain Surface Tuning\u201d method described in Sections II and III to the three-loop autopilot of Figure 1. A schematic of the airframe appears in Figure 4. The autopilot must track a command \u03b3ref in flight path angle by controlling the normal acceleration Az and the pitch rate q. PI control is used for the pitch rate loop and static gains are used for the acceleration and flight path loops. Because the aerodynamic forces and moments vary with the incidence angle \u03b1 and the speed V , the autopilot gains Kp,Ki,Ka,Kg in Figure 1 must be scheduled as a function of \u03b1 and V . For the operating range considered here, \u03b1 varies between -20 and +20 degrees and V varies between 700 to 1400 m/s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000451_taes.2008.4667721-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000451_taes.2008.4667721-Figure3-1.png", + "caption": "Fig. 3. Reference circular curve for impact angle control.", + "texts": [ + " Summing up the above discussions, a circular reference curve is expected to pass the point (x(0),y(0)) with a slope \u00c1(0), to be compatible with a given initial condition, and then the origin with a slope \u00c1de to meet the impact angle specification. Obviously not all conditions can be simultaneously satisfied with a single circle. Therefore temporally let us assume that the initial slope \u00c1(0), equivalently the initial attitude (launching) angle \u00b0(0) =\u00a6\u00a11(\u00c1(0)), is at our disposal which is called the free launching condition. Under this condition the initial slope condition (12) can be safely ignored and remaining conditions uniquely determine a circle C\u00a4ref in Fig. 3 as a reference curve. To be precise the reference curve is only a part (arc) of the circle C\u00a4ref but this difference will not cause confusions as the directional line ` in Fig. 3 indicates the direction of velocity vector ( _x, _y) at impact. This justifies a convention of calling the circle in Fig. 3 as simply a reference circle from now on. As a reference circle C\u00a4ref is defined on the moving frame OXY attached to the target, observers moving with target will see a circular trajectory of a pursuer. With respect to an inertia frame however, the pursuer\u2019s trajectory is a certain type of helix in general. Elementary calculations give that the radius of curvature Rref and the center of the reference circle C\u00a4ref are given by Rref = 2sin(\u00c1de\u00a1\u00be) D \u00af\u0304\u0304\u0304 t=0 \u00b3ref = 1 Rref (\u00a1sin\u00c1de,cos\u00c1de) D := p x2 + y2 (15) where \u00be := arg([\u00a1x \u00a1 y]) 2 (\u00a1\u00bc,\u00bc] (16) denotes the LOS (line-of-sight) angle. REMARK 2 Regarding Rref as a function of time, i.e., Rref(t) in (15), it is a constant function as long as the point (x,y) moves on the circle C\u00a4ref. To show this note that dRref(t) dt = 2W(\u00c1) D2 sin(2\u00be\u00a1\u00c1\u00a1\u00c1de) (17) from the pursuit dynamics (7)\u2014(8) and (16). Also the equality \u00c1\u00a1\u00c1de = 2(\u00bc+\u00be\u00a1\u00c1de) holds while (x,y) moves on the reference circle C\u00a4ref; see Fig. 3. It should also be noticed that Rref(\u00a2) can be negative. Note that the reference circle is not well defined if an initial point (x(0),y(0)) lies on a straight line 1452 IEEE TRANSACTIONS ON AEROSPACE AND ELECTRONIC SYSTEMS VOL. 44, NO. 4 OCTOBER 2008 overlapping the directional line ` of Fig. 3. That is, the following condition \u00be(0) 6= \u00c1de equivalently Rref(0) 6= 0 (18) should hold for a valid reference circle, which is also clear from the radius of the curvature term in (15). In order to clarify the implications of this requirement, let us briefly consider situations where (18) does not hold. Firstly, suppose 0< \u00c1de < \u00bc as in Fig. 3 and \u00be(0) = \u00c1de. Then the reference circle C\u00a4ref becomes a straight line overlapping the line `. As such, if y(0)< 0, then a pursuer is well aligned initially and a perfect impact angle follows with no controls. On the contrary, if y(0)> 0, then a capturing fails. These two special situations can be shown to occur in other cases where \u00c1de = 0, \u00c1de = \u00bc, or \u00a1\u00bc < \u00c1de < 0. Interestingly those extreme cases coincide with the well-known singularities of the pure PNG law neatly characterized in the paper [18]. In this paper the condition (18) is assumed for presentational simplicity. Motivated from [12], we define a set S(\u00c1de)\u00bd R2\u00a3 (\u00a1\u00bc,\u00bc] by S(\u00c1de) := f(p,q,s) : 2arg([p q])\u00a1 s= \u00c1deg (19) for a given \u00c1de 2 (\u00a1\u00bc,\u00bc] where the equality in (19) is defined up to modulus 2\u00bc. Notice that, for a given \u00c1de, the inclusion (x,y,\u00c1) 2 S(\u00c1de) is equivalent to the condition that the point (x,y) moves on the reference circle in Fig. 3. From this fact and the Frenet formulas we have the following result. THEOREM 1 Suppose (x(0),y(0),\u00c1(0)) 2 S(\u00c1de) and the following command, called the curvature control, is employed in the pursuit dynamics (7)\u2014(9): uo = W3(\u00c1) L(\u00c1) Rref = 2 W2(\u00c1) L(\u00c1) _\u00be: (20) Then there exists a finite time tf > 0 such that the solution (x,y,\u00c1) satisfies (i) (x(t),y(t),\u00c1(t)) 2 S(\u00c1de) for all t 2 [0, tf) (21) (ii) lim t!tf (x(t),y(t),\u00c1(t)) = (0,0,\u00c1de): (22) PROOF All claims follow from the Frenet formulas except the last equality (20)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003151_ecce.2013.6647071-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003151_ecce.2013.6647071-Figure2-1.png", + "caption": "Figure 2. Winding configuration without neutral point (NP nonconnected)", + "texts": [ + " Therefore, the prototype motor developed for this paper has the structure in which the stator coil can connect freely and that has 12 output terminals including U1+ and U1\u2013. Six current probes are used to measure the current in the coils, i.e., U1+, U2+, V1+, V2+, W1+, and W2+ at the same time. These outputs are then saved in a data logger. The operating conditions are under no load operation at a frequency of 50 Hz. Figs. 1 and 2 are diagrams of stator winding equivalent circuits: Fig. 1 shows the circuit with the connected neutral point (NP connected) while Fig. 2 shows that with the unconnected (NP non-connected). Both types use the same motor, which can be switched to neutral point on or off. Fig. 3 outlines the static rotor eccentricity where the rotor is displaced vertically toward the U2 coil by 13 % of the normal air-gap length. In addition, the experimental IPM has a maximum erection tolerance of 7 % and hence the total eccentricity value ranges from 13 up to 20 %. III. VOLTAGE DIFFERENCE UNDER ECCENTRICITY This paper introduces a theoretical expression of the circulating current for the purpose of comparing connected and unconnected neutral points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure7.13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure7.13-1.png", + "caption": "Fig. 7.13 Web cam controller", + "texts": [ + " The servomotor is an Hitachi HS-805BB controlled by a PWM (Pulse With Modulation) signal generated by the multi function data acquisition device of the PCI-6025E board. To protect the data acquisition device from extra currents, an electronic circuit has been developed to connect the servomotor to the board. 168 G. Carnevali and G. Buttazzo BookID 175907_ChapID 7_Proof# 1 - 11/4/2009 BookID 175907_ChapID 7_Proof# 1 - 11/4/2009 The control application is developed under the Shark operating systems and consists of three periodic tasks, as shown in Fig. 7.13. The figure illustrates how the different tasks interact to achieve the goal. We have used the semaphores technique to access mutually exclusive resources. Move_Motors is a hard real-time periodic task scheduled using the Earliest Deadline First scheduling algorithm. It implements a PID controller and uses the information stored in the PID parameters buffer to calculate the new position of the motor, which is then stored in the System Status buffer. Msg_Receiver and Msg_Sender are soft real-time periodic tasks scheduled using the Constant Bandwidth Server" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001707_bsn.2010.46-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001707_bsn.2010.46-Figure4-1.png", + "caption": "Fig. 4. Front view of club face and ball at impact showing lie angle and impact location.", + "texts": [ + " Positive z points upward perpendicular to the ground, positive x denotes the intended ball path and positive y points to the golfer while remaining perpendicular to x and z. There are many parameters of a golf swing that affect the trajectory of the golf ball. The goal of our model is to identify parameters that address precision and repeatability of the swing. We focus our attention on the following most critical parameters: \u2022 Face angle (\u03c8) at impact (Figure 2) \u2022 Loft angle (\u03b8) at impact (Figure 3) \u2022 Lie angle (\u03c6) at impact (Figure 4) \u2022 Velocity throughout swing \u2022 Location (x\u2032, y\u2032) of impact on club face (Figure 4) \u2022 Motion path immediately surrounding impact1 \u2022 Tempo: Proportion of back-swing duration to forwardswing duration By examining each of Figures 2, 3 and 4 it can be observed that the most predictable path of travel for the golf ball will occur when loft, face and lie angles are all zero. In fact the motion is so sensitive to error that a \u03c8 = \u00b13 degree face angle will result in an error greater than 15cm for a 3m putt. This may seem minor but the diameter of the target hole is \u2248 11.5cm. One might be tempted to alter the ball path using a non-zero face angle but this practice is discouraged because it is imprecise and difficult to repeat" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003507_amr.819.7-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003507_amr.819.7-Figure2-1.png", + "caption": "Fig. 2 forming principle of RP", + "texts": [ + " Generally, there are two main tool path trajectories for FDM processes: contour parallel path and direction parallel path as shown in Fig. 1. The former focus on the surface quality at the cost of efficiency, and the latter considers the build time reduction while neglects the surface accuracy. Besides the requirements of tool path of general material removal machining, some special requirements also need to be met in the tool path generation of FDM. The forming principle of FDM is displayed in Fig. 2. The path of the nozzle/print head is covered with forming material. The cross section of material in the vertical direction is an approximate semiellipse which is determined by the property of the forming material. The solid boundary of the forming material in the figure is the final surface of one layer. h is the height of the layer and w is the width of the strip of forming material, and a proportional relationship exists between them. If the distance between tool paths is too large, a gap between two adjacent strips of forming material may appear", + " Second, the abrupt corners between line and tiny arc will result in the sharp alternation of speed of nozzle/print head, which can jeopardize the surface smoothness according to Eq. 4. Hence, the original tool paths need some adjustments in the corners to alleviate the influence of the mentioned issues. Tool path adjustment. Changing direction abruptly of the tool path is the root of the occurrence of unfilled areas and unstable speed. At each turn, the filled strip of two adjacent tool paths travelling in different directions might not have enough overlapping rate, a scallop appears between them. Besides, as the forming principle of FDM shown in Fig. 2, an excessive filled area will inherently appear. The Fig.5 displays the two phenomena. The areas \u2460 and \u2461 are excessive filled areas and the area \u2462 is the unfilled area because of the insufficient overlapping rate between the adjacent tool paths. To improve the surface accuracy of sliced layer, the original tool paths at the turn need some modifications to avoid the sharp angle. One feasible approach is shown is Fig. 5. The aim of the modification is increasing the overlapping rate of area \u2462 and decreasing the overlapping rate of area \u2460 and \u2461" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure13.1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure13.1-1.png", + "caption": "Fig. 13.1 A prototype of the RISCbot", + "texts": [ + " The tele-operation tasks are potentially useful in dangerous and unpredictable environments such as at a construction site, space, underwater, service environments, and in nuclear power stations. support from the ground. A mobile manipulator offers a dual advantage of mobility offered by the platform and dexterity offered by the manipulator. For instance, the mobile platform extends the workspace of the manipulator. We are developing and constructing a mobile manipulation platform called RISCbot. The prototype of the RISCbot is shown in Fig. 13.1. Sensor fusion has been an active area of research in the field of computer vision and mobile robotics. Sensor fusion can be defined as a method for conveniently combining and integrating data derived from sensory information provided by various and disparate sensors, in order to obtain the best estimate for a dynamic system\u2019s states and produce a more reliable description of the environment than any sensor individually. Sensor fusion algorithms are useful in low-cost mobile robot applications, where acceptable performance and reliability are desired, given a limited set of inexpensive sensors such as ultrasonic and infrared sensors", + " In [23], the recent Radio Frequency Identification (RFID) was used to improve the localization of mobile robots. This research studied the problem of localizing RFID tags with a mobile robot that is equipped with a pair of RFID antennas. Furthermore, a probabilistic measurement model for RFID readers was presented in order to accurately localize RFID tags in the environment. We are developing and constructing the mobile manipulation platform called RISCbot (the prototype of the RISCbot is shown in Fig. 13.1). The RISCbot mobile manipulator has been designed to support our research in algorithms and control for autonomous mobile manipulator. The objective is to build a hardware platform with redundant kinematic degrees of freedom, a comprehensive sensor suite, and significant end-effector capabilities for manipulation. The RISCbot platform differs from any related robotic platforms because its mobile platform is a wheelchair base. Thus, the RISCbot has the advantages of the wheelchair such as high payload, high speed motor package (the top speed of the wheelchair is 6 mph), Active-Trac and rear caster suspension for outstanding outdoor performance, and adjustable front anti-tips to meet terrain challenges" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003830_icmech.2013.6518544-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003830_icmech.2013.6518544-Figure1-1.png", + "caption": "Fig. 1 The kinematic arrangement of SURALP", + "texts": [ + " They can be separated based on the frequency band using two low pass filters (LPF) with different frequencies. The offset error is the output of the low frequency LPF. This offset error is subtracted from the output of the higher frequency LPF to obtain the acceleration error [11, 12]. III. SIMULATION RESULTS The simulations are carried on 12 degrees of freedom (DOF) biped model. It consists of two 6-DOF legs and a trunk connecting them. Three joint axes are positioned at the hip, two joints are at the ankle and one at the knee (Fig. 1). The numerical values of the parameters (Table ) are taken to match our experimental humanoid robot SURALP (Sabanci University Robotics Research Laboratory Platform) [16]. The details of contact modeling and simulation algorithm are in [17]. The modeled reaction forces suffer from peaks, so that Kalman filter is used for smoothing the modeled reaction forces. The coordinate frames are shown in Fig. 2. All the measurements and calculation are in the world frame. The transformation is done using the rotational matrix obtained by the author in [18]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000277_s12283-008-0002-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000277_s12283-008-0002-3-Figure1-1.png", + "caption": "Fig. 1 Throwing arm dynamic model. sR; sP; sY : the joint axes of the upper trunk posterior/anterior lean, medial/lateral lean, counterclockwise/clockwise rotation. s1, s2, s3: the joint axes of the shoulder horizontal adduction/abduction, adduction/abduction, external/internal rotation. s4: the joint axis of the elbow extension/flexion. s5, s6, s7: the joint axes of the wrist ulnar/radial flexion, flexion/extension, supination/pronation. lt: the vectors pointing from the mid-point of the shoulder line to the throwing shoulder. lu, lf, lh: the vectors pointing from the proximal end to the distal end of the upper arm, forearm and hand. R0: global reference frame. R1, R2, R3, R4: segment reference frames at the upper trunk, upper arm, forearm and hand. The X, Y and Z components in R1, R2, R3 and R4 were determined by the previous studies (see text)", + "texts": [ + " In the present study, to consider the contribution to standard baseball pitching movements, trials in which the ball speed exceeded 120 km/h (about 33.3 m/s) and crossed the strike zone at home plate were analysed. Two pitches from each of the three pitchers (total six) satisfied the two conditions. 2.2 The throwing arm dynamic model The throwing system was modelled as a linked system consisting four rigid bodies (upper trunk, throwing upper arm, forearm and hand) and the ball. The model involved the upper trunk rotations about three axes and the throwing shoulder, elbow and wrist joints including 7 degrees of freedom (Fig. 1). Each joint angle of the model was calculated using the Eulerian angle convention described by Chao [4]. The Eulerian convention is defined as the rotational sequence of motion that follows the order of coordinate axes Z, y0 and x00, and the Eulerian angles q1, q2 and q3 represent as rotational angles about Z, y0 and x00 of the right-handed reference frame, shown in Fig. 2. For the upper trunk, the counterclockwise (+) and clockwise (-) rotation hY, medial (+) and lateral (-) lean hP, and posterior (+) and anterior (-) lean hR were defined as the Eulerian angles q1, q2 and q3 about Z, y0 and x00, respectively", + " For the wrist, the angles of the ulnar (+) and radial (-) flexion h5, flexion (+) and extension (-) h6, and supination (+) and pronation (-) h7 were defined as the Eulerian angles q1, q2 and q3 about Z, y0 and x00, respectively. The neutral position of the hand reference frame was coincident with the forearm reference frame in three directions. In this case, the unit vectors corresponding to the directions of Z, y0 and x00 were defined as the joint axes of the ulnar/radial flexion (s5), flexion/extension (s6), and supination/pronation (s7), respectively. sR, sP, sY and sj (j = 1\u20137) described in Fig. 1 are defined as the vectors relative to the global reference frame. The throwing arm dynamic model was used to calculate the kinematic and kinetic variables of the upper trunk and throwing arm. The angular velocity \u00f0xt\u00de and acceleration \u00f0 _xt\u00de of the upper trunk relative to the global reference frame were obtained as follows: xt \u00bc sY _hY \u00fe sP _hP \u00fe sR _hR \u00f01\u00de _xt \u00bc sY \u20achY \u00fe sP \u20achP \u00fe sR \u20achR \u00fe sY _hY sP _hP \u00fe \u00f0sP _hP \u00fe sY _hY\u00de sR _hR: \u00f02\u00de The upper arm angular velocities \u00f0x1;x2 and x3\u00de and accelerations \u00f0 _x1; _x2 and _x3\u00de about sj (j = 1\u20133), the forearm angular velocity \u00f0x4\u00de and acceleration \u00f0 _x4\u00de about s4 , and hand angular velocities \u00f0x5;x6 and x7\u00de and accelerations \u00f0 _x5; _x6 and _x7\u00de about sj (j = 5\u20137) were obtained as follows: xj \u00bc xj 1 \u00fe sj _hj \u00f0j \u00bc 1 7\u00de \u00f03\u00de _xj \u00bc _xj 1 \u00fe sj \u20achj \u00fe xj sj _hj \u00f0j \u00bc 1 7\u00de: \u00f04\u00de Here x0 (in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure8-1.png", + "caption": "Fig. 8 Configuration 5 of the example", + "texts": [ + " 7 a , the revolute pair between Links 4 and 5 is then frozen. The mechanism is termed as Configuration 4 and is shown in Fig. 7 b . If this configuration comes from Configuration 3, the adjacency matrix A4 can be obtained by multiplying \u22121 on A3 3,4 , A3 4,3 , A3 4,5 , and A3 5,4 , i.e., A4 = 0 1 0 0 1 1 0 1 0 0 0 1 0 1 0 0 0 1 0 \u2212 1 1 0 0 \u2212 1 0 11 ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash 5 If Link 1 is attached together with Link 5 by a pin, P, as shown in Fig. 8 a , the revolute pair between Links 1 and 5 is frozen. The mechanism is termed as Configuration 5 and is shown in Fig. 8 b . If this configuration comes from Configuration 4, the adjacency matrix A5 can be obtained by multiplying \u22121 on A4 4,5 , A4 5,4 , A4 1,5 , and A4 5,1 , i.e., A5 = 0 1 0 0 \u2212 1 1 0 1 0 0 0 1 0 1 0 0 0 1 0 1 \u2212 1 0 0 1 0 12 Note that among all possible configurations, any one can be transformed to any other one. Though, the adjacency matrix can always catch the process of the change. Figure 9 shows such an example. This is a planar metamorphic mechanism with five links. There is a spring embedded in Link 1, which can push Slider 2 moving along the slot" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003237_s11665-013-0583-2-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003237_s11665-013-0583-2-Figure4-1.png", + "caption": "Fig. 4 Typical: (a) geometrical (after reflection across the weld boundary); and (b) meshed models used in the thermomechanical GMAW process module", + "texts": [ + " In the remainder of this section, a brief overview is provided by the key aspects of the thermomechanical GMAW process module, such as: (a) geometrical model; (b) meshed model; (c) computational algorithm; (d) initial conditions; (e) boundary conditions; (f) mesh sensitivity; (g) material model; and (h) computational tool. Journal of Materials Engineering and Performance 2.3.1 Geometrical Model. The computational domain comprising two workpieces to be butt-welded initially (i.e., before welding) possessed a rectangular-parallelepiped shape with the following dimensions: (60 mm9 60 mm9 10 mm). The axes of the parallelepiped are aligned with the global x-y-z Cartesian system, as indicated in Fig. 4(a). The following orientation of the computational domain is chosen: (i) the weld contact interface is set orthogonal to the x-axis; (ii) the weld gun travel direction is aligned with the y-axis; while (iii) the workpiece through-the-thickness upward normal direction is aligned with the z-axis. The origin of the coordinate system (x = y = z = 0) is placed at the mid-value of x, minimum value of y, and at the minimum value of z of the computational domain. To geometrically model the FZ, a different approach was adopted from that used in Ref 5, 6", + " To account for the filler-material deposited into the weld-pool and the resulting formation of the weld bead, the top surface of the workpiece was dynamically reshaped in accordance with the advancement of the weld gun. The profile of the resulting weld bead is consistent with the spraying-rate normal distribution function, as discussed in section 2.2. The resulting geometry of the two workpieces to be butt-welded, at the instant when the weld gun has traveled half of the distance in the y-direction, is depicted in Fig. 4(a). It should be noted that the computational domain described above is symmetric about x = 0 and, hence, only one (right, when looking along the direction of motion of the welding-gun, in this study) half of this computational domain had to be explicitly analyzed. 2.3.2 Meshed Model. The selected half of the computational domain is meshed using between 16,425 and 52,429 eight-node, first-order, thermomechanically coupled, reducedintegration, hexahedral continuum elements. Figure 4(b) shows a close-up of the typical meshed model used in this study. Examination of this figure clearly reveals the presence of the deposited weld bead. 2.3.3 Computational Algorithm. All the calculations carried out within the present module are based on a transient, fully coupled, thermomechanical, unconditionally stable, implicit finite-element algorithm. At the beginning of the analysis, the computational domain is supported over its bottom (z = 0) face, made stress free, and placed at the ambient temperature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001203_j.na.2009.01.238-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001203_j.na.2009.01.238-Figure2-1.png", + "caption": "Fig. 2. Schematic view of rotor-AMB system with coordinates that describes its motion.", + "texts": [ + " For simplicity only the force acting on vertical direction is given. Horizontal force can easily be derived with the same procedure as described above. In this section dynamical equations of a rigid rotor supported by two AMBs are derived. Magnetic bearings levitate the rotor with variable stiffness and damping in each axis separately. When the rotor approaches the bearing stator, AMB produces a net force in the opposite direction to move the rotor to the bearing rotation centre. The rotor-AMB system is depicted in Fig. 2 where this rigid rotor with a mass unbalance rotates about z-axis with an angular velocity of \u03c9. Similar to the mathematical model given in [11], equations of motion of the rigid rotor with an unbalanced disk can be derived by using Newton\u2013Euler formulae where \u2211 F = ma and \u2211 M = I\u03b1. Mx\u0308G = Fx1 + Fx2 + Fxu +Mg (5) Ix\u03b8\u0308y = \u03c9Iz \u03b8\u0307x + Fx1L1 \u2212 Fx2L2 (6) My\u0308G = Fy1 + Fy2 + Fyu (7) Iy\u03b8\u0308x = \u2212\u03c9Iz \u03b8\u0307y \u2212 Fy1L1 + Fy2L2, (8) where Ix and Iy are transverse mass moment of inertia about x- and y-axes respectively, Iz is polar moment of inertia about z-axis,M denotes mass of rotor, L1 and L2 are the distances between the centre of mass and magnetic bearings on the lefthand side and on the right-hand side respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001797_aqtr.2010.5520862-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001797_aqtr.2010.5520862-Figure3-1.png", + "caption": "Figure 3. IR sensor locations.", + "texts": [ + " This prevents IR interference from common sources such as sunlight and indoor lighting. The key to making each IR LED/detector pair work is to send 1 ms of 38.5 kHz frequency signal and then, immediately store the IR detector\u2019s output in a variable. The IR detector\u2019s output state when it sees no IR signal is high. When the IR detector sees the 38500 Hz harmonic reflected by an object, its output is low. The sensors are mounted on the front platform at 60\u00b0 one each other from the central axis, like in Fig. 3. The sensors include the left side, front and right side locations of the robot, namely Left_Sensor (SL), Front_Sensor (SF), Right_Sensor (SR). The detection distance for each sensor is more than 10 cm with 80% sensitivity. After some tests we concluded that IR sensors have a detection range of 10 cm around the mobile robot platform and the detection angle is more about 900 in the front of the robot, like in Fig. 4. The mechanical structure handling this architecture is based on the two DC motors controlling through gears two differential wheels" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000403_0005-2736(80)90202-3-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000403_0005-2736(80)90202-3-Figure9-1.png", + "caption": "Fig. 9. p H d e p e n d e n c e o f t he ze t a p o t e n t i a l o f p r o t o p l a s t s (o) a n d o f i n t a c t y e a s t cells (o). E a c h point r e p r e s e n t s t h e m e a n va lue o f a t l eas t 20 d e t e r m i n a t i o n s .", + "texts": [ + "5 mM, by protoplasts and also by yeast cells was linear for at least 10 min. The rate of uptake was not affected by enzymic removal of the cell wall, in agreement with the results of the Rb \u00f7 uptake experiments described above. Plots of K \u00f7 uptake by intact yeast cells showed an intercept with the ordinate (Fig. 8 is shown as an example), probably originating from accumulation of K \u00f7 in the cell wall space. The pH dependence of the electrophoretic mobility of yeast cells and that of protoplasts did not differ significantly (Fig. 9). Qualitatively, a similar pH dependence of the electrophoretic mobility has been found also with other strains of yeast [ 15]. Typically, the isoelectric points were rather low (less than pH 3). The present results show that under the experimental conditions applied, yeast cells were completely converted into protoplasts by the enzymic treatment. After 30 min of enzymic treatment the cells became completely osmofragile (Fig. 1), the ovoid shape of the yeast cell became spherical (Figs. 2 and 3) and in freeze-etch replicas of cell preparations characteristic invaginations of the plasma membrane could be visualized (Fig", + " Results of experiments performed with the K\u00f7-sensitive electrode support the conclusion that the transport properties of the cells were not affected by the enzymic t reatment (see Fig. 8). The rates of K \u00f7 uptake were identical for the protoplasts and for the intact yeast cells. We argued in a series of earlier publications [20--22] that the K m of ion translocation across cellular membranes depends on their surface potential. Our observation of a single relationship between the Km of the Rb \u00f7 uptake and the zeta potential of the yeast cell [23] supports that notion. Fig. 9 shows that the zeta potential of yeast cells, over a large pH range, was not affected significantly by enzymic removal of the cell wail. Similar results were found in a comparative s tudy of the zeta potential of intact yeast cells and yeast plasma membrane vesicles [12]. Also as no difference in K m of Rb \u00f7 uptake was observed between intact yeast cells and protoplasts it might be hypothesized that the zeta potential of protoplasts and of intact yeast cells, as well, is mainly determined by charges on the plasma membrane, and that charges in the cell wall are of minor importance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002628_amm.373-375.38-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002628_amm.373-375.38-Figure2-1.png", + "caption": "Fig. 2. The dynamic model of the stiff rotor on elastic supports with the self-balancer of a pendulum type", + "texts": [ + " The stroboscopic tachometer was used to supervise the location of pendulums and measurement of angular speeds of the rotor and pendulums 12. The mass and the moments of the rotor inertia, the stiffness of its supports, and the moments of the pendulums inertia are selected in such a way as to provide stability of the self-balancing motion mode. In particular, that is the condition according to which the angular speed of the rotor rotation has to be more than its critical speeds [9]. The dynamic model in the form of the stiff horizontal rotor mounted on elastic isotropic supports (Fig. 2) was chosen for the analytical research. While formulating the equations of motion of the model for case of acceleration and in the established mode the following generalized coordinates were selected: y, z are motions of 01 point from the position of static balance of the rotor in the direction of axes 0y and 0z (01 is a point of intersection of the rotor axis with the plane passing through its center of masses perpendicularly to the axis); \u03b8, \u03c8 are the angles between the axis \u0445 and the projections of the rotor axis to the \u0445\u0443 and \u0445z coordinate planes; \u03c6 is the rotation angle of the rotor about its axis; \u03c61, \u03c62, \u03c63, \u03c64 are the rotation angles of the pendulums" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000465_iet-smt:20060018-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000465_iet-smt:20060018-Figure4-1.png", + "caption": "Fig. 4 Magnetic forces in four-pole machine", + "texts": [ + " Inspection of the results for mode 4 in Table 2 leads to the conclusion that, for this mode, the contributions of the magnetic forces and the magnetostriction to the vibrations are in counter phase and thus are subtracted. As an illustration, the calculated magnetic forces and the calculated magnetostriction in a quarter of the four-pole machine are depicted in Figs. 4 and 5, respectively. The black arrows represent the nodal forces, and the white lines represent the magnetic flux lines. The nodal forces in Fig. 4 have a larger magnification factor than the nodal forces in Fig. 5. Application of the same magnification factor for both figures would lead to very small nodal forces in Fig. 4, and so, at first sight, it would appear that the magnetic forces are much smaller than the forces due to the magnetostriction. However, the finite element mesh has more elements near the edges of the teeth than in other parts of the stator. These elements logically have a smaller surface, and so it follows that, in these regions, the nodal forces are smaller (see (3)), but greater in quantity. IET Sci. Meas. Technol., Vol. 1, No. 1, January 2007 23 In this paper, a numerical method has been elaborated to calculate the vibrations in magnetised electrical steel caused by magnetic forces and magnetostriction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000219_iecon.2008.4758458-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000219_iecon.2008.4758458-Figure2-1.png", + "caption": "Fig. 2. Schematic of an n-phase, two-pole PM machine under phase-1 open circuit fault: (a) rotor position \u03b1, and (b) rotor position \u2013 \u03b1.", + "texts": [ + " It is found that the optimal fault-tolerant control of the five-phase PM machine under single phase open-circuit condition can be achieved by using only fundamental sinusoidal currents. The derivation of the solution is presented in this work. The implementation of the proposed fault-tolerant control technique is discussed. Finite Element simulation results and experimental results of a five phase PM machine are presented to substantiate the proposed solution. It is found that the machine can successfully continue its operation with satisfactory performance. Consider a two pole, n-phase, PM machine as shown in Fig. 2, where phase-1 is under open circuit fault. Assume two stator phases, phase-k and phase-m, are located symmetrically in space with respect to the magnetic axis of faulty phase-1 (Fig 2). Phase-1, phase-k, and phase-m are located at angles: 0, nk.2\u03c0 , and nk.2\u03c0\u2212 , respectively. Further consider Fig 2(a), when the rotor PM axis is located at an angle\u03b1 . Assume, at this rotor position to produce a certain amount of output torque, phase-k and phase-m winding should be excited with currents 1ki and 1mi , respectively. Further consider Fig 2(b), where the rotor PM axis is positioned at an angle \u03b1\u2212 . Assume, at this position, to produce the same output torque, the currents of phase-k and phase-m are 2ki and 2mi , respectively. From the symmetry of rotary machine structure and from the mirror symmetry of the spatial location of phase-k and phase-m with respect to phase-a axis it can be perceived that: under phase-a open-circuit fault, the phase-k current at rotor position \u03b1 should be equal to the phase-m current at rotor position \u03b1\u2212 and vice versa ( 1221 , mkmk iiii == )" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001368_j.mechatronics.2010.09.010-Figure26-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001368_j.mechatronics.2010.09.010-Figure26-1.png", + "caption": "Fig. 26. A simplified model of the actuator system.", + "texts": [ + " The harmonic drive has the advantage of no backlash, high gear ratio and high torque transfer capability within a small volume, thus yielding a flat, compact surface mount actuator. The electrical equation of the actuator motor is given by: ei \u00bc iaRa \u00fe La d dt \u00f0ia\u00de \u00fe ka _#m; tion of the actuator. 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 \u22124 \u22123 \u22122 \u22121 0 1 2 3 4 Square wave tracking response Time (sec) de g Commanded Actual where ei is the control voltage applied to the motor, ia is the motor armature current, and Ra and La are the motor armature resistance and winding inductance, respectively (see Fig. 26). ka is the motor where Tm is the motor torque, and Jm and Bm are the moment of inertia of the motor armature plus load, and the viscous damping coefficient of the rotating assembly. The motor torque Tm is directly proportional to the armature current: Tm = kaia. The rotation of the actuator output shaft # is related to the motor shaft rotation by the overall gear ratio or deceleration constant n:# = #m/n. The nonlinearities in the actuator arise primarily from the voltage and current limitation in the control and drive circuits, and the friction of the moving parts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003082_j.optlaseng.2013.10.010-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003082_j.optlaseng.2013.10.010-Figure2-1.png", + "caption": "Fig. 2. Transfer modes, to the left: desired transfer with d\u00bc dnom . In the center: dodnom causes stubbing. To the right: d4dnom , with risk of droplet formation. Stick-out length, l, is indicated in leftmost illustration. (For interpretation of the references to color in this figure caption, the reader is referred to the web version of this article.)", + "texts": [ + " This control has previously been carried out through 3D-scanning of the deposit before and after each deposited layer and subsequent adjustments of the deposition path [16]. Even as the challenges for this wire based method are substantial, so are the benefits. There is almost a 100% material usage and there is no need for handling of excess powder. Also, for simpler and larger geometries, LMD-w enables higher deposition rates and might also give better surface finish and material quality than powder based LMD [8,17]. The geometry of the LMD-w weld pool, especially viewed from the side, is affected by the distance from the wire tip to the substrate as seen in Fig. 2 [8]. If a change in pool geometry due to a change in distance between the tool and the workpiece results in a measurable change in electrical resistance, it would be possible to monitor said distance by resistance measurements. This idea and its implementation for control purposes is the main contribution of this work. The prospects of this idea lie in its simplicity and ease of use. Monitoring voltage across, and current through, the weld pool and the wire is relatively simple from an implementation perspective", + " The line scanner, apart from being an additional instrument bringing costs and implementation efforts, takes time from the deposition itself since scanning and processing are not carried out in parallel. In this paper, the control strategy is introduced, followed by an empirical model of weld pool and wire resistance followed by the controller design. The implemented control system's performance is evaluated by comparison with an uncontrolled system. Finally, the possibilities of improvement are briefly discussed. As concluded by Herali\u0107 et al. [16], the distance between the tool and the workpiece, d, affects the weld pool geometry as indicated in Fig. 2. If d gets too large, the material transfer link between the wire and the pool, indicated in red in the figure, is broken. A weak but sustained link does not typically present a problem. However, it indicates that the limit of the process window is nearby and that d should not be increased any further. If the material transfer link is broken, the weld pool will be sustained by the laser, while a droplet will form at the tip of the wire. This droplet will eventually grow enough to get pulled down by gravity into the weld pool", + " All deposition was done in the form of single beads on a 3 mm thick Alloy 718 plate. In order to protect from oxidation, an argon filled processing chamber was used. The wire feed rate was kept constant at the nominal wire feed rate, vnom, except during excitation for creating the resistance-distance model. Welding current, traverse speed and laser power were kept constant throughout the experiments. The resistance, R, of the wire and the weld pool is the sum of the resistance of the wire, Rw, mainly dependent on wire stick-out length, l, as indicated in Fig. 2, resistance in the weld pool, Rp, and circuit resistance, Rc , which is constant and originates in contacts and cables, etc. R\u00bc Rw\u00feRp\u00feRc \u00f02\u00de By values of resistivity, \u03c1, for the used material, and the assumption that d affects l, Rw can be related to d since Rw is a function of l. The resistivity of Alloy 718 increases with temperature [20]. Estimating the contribution of Rw to total variation in R using a resistivity value for room temperature thus gives a lower limit since wire resistivity typically will exceed that of room temperature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002779_j.phpro.2013.03.104-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002779_j.phpro.2013.03.104-Figure1-1.png", + "caption": "Fig. 1. Dimensions of the section to model (left), in mm, and the constructed mesh for CFD modelling (right)", + "texts": [ + " Using the models, the fluid flow and thermal characteristics in laser welding of an aluminium alloy (AA5083) have been studied. In combination with corresponding experimental results, the numerical results have been used to better understand the porosity formation mechanisms in welds, and to study the relationship between porosity levels in welds and laser welding parameters. A CFD model was established with Ansys13 Fluent to simulate the laser welding of the aluminium alloy AA5083. The dimensions of the workpiece and the computational mesh are shown in Figure 1. The moving laser beam irradiated the top surface of the workpiece. A three-dimensional mesh was constructed as shown in Figure 1, which has 160000 elements, 173061 nodes, with a minimum element size of 0.033\u00d70.033\u00d70.2mm3. During laser welding, relative motion between a laser beam and a workpiece results in the formation of a molten pool, which then solidifies to form a weld. To model this process, it was assumed that the laser beam (the heat source) and the mesh coordinates were fixed, and the workpiece was moved in the negative xdirection at a given welding speed. The other assumptions made in numerical model included that: The fluid flow in the weld pool was transient, laminar and incompressible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000225_001-Figure13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000225_001-Figure13-1.png", + "caption": "Figure 13 The outline of a practical pC02 electrode, suitable for measuring the pC0z of a blood sample, The electrode is housed in a thermostatically controlled holder (not shown) with the membrane adjacent to the blood sample chamber (not shown).", + "texts": [ + "00 for a perfectly well behaved electrode, when the NaHC03 concentration was above 10-3 M, thus validating the theoretical prediction. Thus if the S of an electrode is - 1.00, for each change in pCO2 by a factor of 10, there is a unit change in the pH of the bicarbonate solution. At a steady temperature of 3 7 T , this would be registered as a voltage change of 61-5 mV. In practice, most electrodes have an S close to - 1.00, and the range is typically - 0.93 to - 0.98. 4.1.2 Practical design The practical pCO2 electrode is illustrated in figure 13. The end of the pH glass electrode is covered by a matrix of nylon mesh, Joseph paper or Cellophane, which holds a thin film of solution containing 0.1 M NaHC03 with some NaCl and sometimes silver chloride. An Ag/AgCl reference electrode, in physical contact with the bicarbonate solution, completes the electrical circuit to the pH electrode. The potential difference between the pH electrode and the reference electrode is registered by a high impedance (loll Q) voltmeter. The pH electrode is separated from the test solution by a Teflon membrane permeable to CO2 molecules" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003889_isie.2013.6563767-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003889_isie.2013.6563767-Figure1-1.png", + "caption": "Fig. 1 wheeled mobile robot", + "texts": [ + " A rotating reflector is added in LRF, such that the detection range can be expanded from point to plane. A local map is constructed by the information obtained from the LRF; however, the disadvantage is that the LRF\u2019s cost is expensive. To choose an appropriate laser sensor while according to user\u2019s requirement is very important. This paper presents PSO and Fuzzy control algorithm for the WMR to explore in the unknown environment. II. ROBOT SYSTEM STRUCTURE The wheeled mobile robot is shown in Fig. 1. It has two DC motors and has encoder in each one. The encoders are used to detect how far the WMR has moved. The LRF was installed on the top of the WMR, and can be connected via USB interface through a notebook. Fig. 2 shows the LRF, the product name is URG-04L), the scan angle is 240 degree, and angular resolution is 0.36 degree, max detection distance is 4000mm. The sensing principle is similar to the sonar sensor, but actually it\u2019s using laser light instead of sound to create two dimensional maps of the proximity to nearby objects" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001842_med.2009.5164715-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001842_med.2009.5164715-Figure2-1.png", + "caption": "Fig. 2. Quadrotor platform. The exterior frame protects the propellers in case of collisions during experiments.", + "texts": [ + " Except for hover, the expression for the rotor wash induced velocities can not be obtained in closed-form, creating difficulties when the model is used to design certain types of controllers. The approach taken in this paper is to model only the most important elements of the quadrotor that define its behavior at hover and ignore the ones that have a significant effect only at high speeds. The derivation of the nonlinear dynamics is performed in the North-East-Down (NED) inertial coordinates and in the x-y-z body-fixed coordinates (Fig. 2). Variables resolved to the inertial axes will be denoted by an e subscript and the ones resolved to the body axes will have the b subscript. The attitude is represented using quaternions. They parameterize the rotation from the inertial reference frame to the body frame using four values. The first is a scalar and the rest form a vector: ( ) ( ) \u22c5 = = r q 2sin 2cos 3 2 1 0 \u03b1 \u03b1 q q q q , (1) where \u03b1 is the angle of rotation and r is the axis around which the rotation is made, with the three components resolved to the inertial axes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure2-1.png", + "caption": "Fig. 2 Schematic representation of generation of a spiral bevel gear", + "texts": [ + " 1 a , Ao is the outer cone distance Fig. 1 a , and Ru is the mean cutter radius. Each type of taper has its recommended interval for the mean cutter radius Ru as given in Ref. 19 and is represented as well in Table 1. All those variables that are affected by the dedendum angle will depend obviously on the type of taper. Basic gear machine-tool settings are represented by those variables that define the geometry of the head-cutter, its position on the cradle, and their position in the cutting machine. Figure 2 shows schematically the head-cutter and a spiral bevel gear. Figure 3 a shows the blade of the head cutter with two edges that allow the generation of convex and concave sides of the gear teeth and Fig. 3 b shows the coordinate systems applied for gear generation. Here, coordinate system Sg is rigidly connected to the head-cutter whereas coordinate system Sc2 is rigidly connected to the cradle, system Sm2 is the system wherein the rotation of the cradle is defined, systems Sa2 and Sb2 are auxiliary coordinate systems to allow the gear blank to be installed, and system S2 is rigidly connected to the gear blank" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003551_iros.2011.6094417-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003551_iros.2011.6094417-Figure4-1.png", + "caption": "Fig. 4: Definition of pan and tilt angles", + "texts": [ + " The visual input data are processed by an adaptive Kalman filter that estimates the actual absolute Cartesian position pf t (k) of the target at sampling time t = kT and predicts its next position pf t (k + 1) at time t = (k + 1)T , both expressed with reference to the head base frame. The kinematic redundancy of the robot head w.r.t. the pointing task is exploited using a weighted pseudoinverse of the task Jacobian, and introducing robot self-motions in the null space of the task Jacobian. The main components of this control concept are illustrated next. The gaze direction of the robot head is a unit vector g from the origin of the end-effector (camera) frame and pointing to the target. With reference to Fig. 4, this direction is characterized by a two-dimensional vector of pointing angles \u03b8 = (\u03b8pan \u03b8tilt) T \u2208 (\u2212\u03c0/2, \u03c0/2)\u00d7 (\u2212\u03c0/2, \u03c0/2) as g = \u239b \u239d gx gy gz \u239e \u23a0 = \u239b \u239d cos \u03b8pan cos \u03b8tilt sin \u03b8pan cos \u03b8tilt sin \u03b8tilt \u239e \u23a0 . (1) Excluding the boundary values \u00b1\u03c0/2 of the tilt angle, the adopted description does not suffer any singularity. The pointing angles are evaluated as: \u03b8pan = atan ( gy gx ) , \u03b8tilt = atan \u239b \u239d gz\u221a g2x + g2y \u239e \u23a0 . (2) The pointing direction n of the robot head is chosen as the x axis of the end-effector (camera) frame, and is determined once the kinematic model of the head is established" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001250_iecon.2009.5415267-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001250_iecon.2009.5415267-Figure3-1.png", + "caption": "Fig. 3. Schematic of Variable Pitch Propeller (VPP) System", + "texts": [ + " Therefore, using (5) and (8), the rotational dynamics for the roll angle can be represented by: \u03c6\u0308 = /Jx (11) where, the sum of moments = F \u00b7d\u2212C \u03c6\u0307 \u03c6\u0307 and F = f1\u2212f2 is the force difference between the right and left rotor and d is the distance from the center of mass to each rotor. Coefficient C \u03c6\u0307 = 0.36 is known as the roll damping derivative. Then (11) can be rewritten as follows: and can be calculated using the following expression: T = Ct\u03c1n 2D4 where Ct is the thrust coefficient, \u03c1 is the density of the air, n is the number of revolutions per second of the motor and D is the diameter of the propellers. The thrust coefficient is 978-1-4244-4649-0/09/$25.00 \u00a92009 IEEE 2343 a function of the pitch angle propeller \u03d5, which is shown in Figure 3. The thrust coefficient in a linear region is given by: Ct = Ct\u03d5\u03d5 where Ct\u03d5 = 0.0025 represents the thrust slope with respect to the VPP angle. This value was estimated for an operational range 5\u25e6 \u2264 \u03d5 \u2264 15\u25e6. In Figure 3, it can be seen that the aerodynamic pitch moment of the blades must be equal to the moment generated by the servo mechanism. Considering that the blade profile corresponds to the NACA0014, the following approximation can be used to obtain the blade pitch moment: mb = \u03c1V 2 tb Sbcb 2Jyb [ Cm\u03d5 \u03d5+ Cm\u03d5\u0307 \u03d5\u0307 ] = ksfs\u03b4s (12) where the subscript b denotes the blade,Cm\u03d5 = \u22120.0019, is the estimated blade pitch moment coefficient slope with respect to \u03d5 and Cm\u03d5\u0307 = 1.6 \u00d7 10\u22125 is a stability derivative generated by the variation of the VPP rate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001117_045104-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001117_045104-Figure10-1.png", + "caption": "Figure 10. Mathematical model of surface of the tip ball of the probe and measured surface.", + "texts": [ + " (e) Sensing error of the displacement sensor in the probe head. The displacement sensor of the probe accommodates the sensing error Gerr, and the input displacement to the probe is not identical to the output measured one. The relationship between the input din and output dout is expressed using the following equation: dout = (1 + Gerr)din. (4) The probe and the measured object are formed in the threedimensional space in the VGC. The gear measurement is simulated by solving the contact problem of the probe and the measured surface. Figure 10 shows the model of the probe and the measured object. The probe position vector pNC of the numerical control command is defined using the following equation, where the x, y, z-axial command positions are xNC, yNC and zNC: pNC = \u239b \u239d xNC yNC zNC \u239e \u23a0 . (5) The vector of the position error of the tip ball of the probe is defined using the following equation: pp = \u239b \u239d Xp Yp Zp \u239e \u23a0 . (6) The y-axial displacement vector of the probe due to the form deviation of the measured object is defined using the following equation: d = \u239b \u239d 0 din 0 \u239e \u23a0 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001181_1.2983146-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001181_1.2983146-Figure8-1.png", + "caption": "Fig. 8. a At t=100 ms there is a net force F=20 N on the bat plus a couple C=10 Nm. F and C act on the handle near the left end of the bat. The knob at the far left end helps to prevent the bat from slipping out of the batter\u2019s hands. b The force exerted by each arm, assuming that the left arm exerts the small force component in a direction parallel to the long axis of the bat.", + "texts": [ + " However, the bat center of mass moves in the opposite direction to the handle end during the initial part of the swing, so the net transverse force on the handle must act in the opposite direction to the direction of motion of the 39Rod Cross icense or copyright; see http://ajp.aapt.org/authors/copyright_permission handle. Both arms therefore pull in opposite directions on the handle, the right arm exerting a greater pull force than the left arm. The magnitude and direction of the transverse force exerted by each arm can be estimated by combining information on the net force, the applied couple, and the axis of rotation of the bat. Consider the situation in Fig. 8 a , which shows the forces exerted by the two arms at t=100 ms, early in the swing. The net force on the bat is 20 N and it acts at an angle =60\u00b0 to the long axis of the bat. In addition, the two arms apply a couple C=10 Nm, which can be represented by equal and opposite 100 N forces acting at right angles to the bat and spaced 0.1 m apart. As a first approximation, each arm therefore exerts an equal and opposite transverse force of about 100 N on the bat. This estimate can be improved by considering the location of the rotation axis of the bat", + " The net force component acting in a direction parallel to the long axis of the bat does not contribute to the torque acting on the bat. We can assume that the net transverse force and the torque on the bat arise from two oppositely directed force components F1 and F2, acting perpendicular to the bat and spaced a distance 0.1 m apart. In that case the net force F2\u2212F1=MdV /dt, where dV /dt=Ld /dt. Because the net torque is equal to Icmd /dt, it is easy to show that F2 /F1=1.18 when L=0.3 m and Icm =0.039 kg m2. The force exerted by each arm must therefore be approximately as shown in Fig. 8 b . The only uncertainty is whether it is the left or the right arm that supplies the small force component acting parallel to the long axis of the bat, or whether each arm contributes about equally. A similar analysis can be applied later in the swing. For example, at t=340 ms the net force on the bat is 300 N acting at =\u221220\u00b0 to the long axis of the bat, the main component being due to the centripetal force. In this case the torque due to F acts in the correct direction to increase the angular velocity of the bat, but it is opposed by a negative couple, 40 Am" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001557_med.2009.5164716-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001557_med.2009.5164716-Figure4-1.png", + "caption": "Fig. 4. Aircraft construction with respect to horizontal airflow", + "texts": [ + " In case of torque equations the angles between the forces and directions are easily derived from basic geometric relations which results in the elemental torque equations: Er% E> ( sEkE> s U* \u00b7 ,lm Ut ( 5 6W ( 8W (8) Er , E> ( \u221a sEFE>s v* ( 8 \u00b7 ,lm Ut ( 5 6W w 8 \u00b7 cde Ut ( 5 6Wx (9) Using the same methods which were used for force calculation the following momentum coefficients were calculated: ,r T+yz Q ,q ( ,_ V 0 ] w *V,lm U56 9W n{ n| } z (10) ,r * \u221a ,_ w V \u00b7 cde 9 T 4~0 ( + \u0398 \u03bb ] (11) ,r * \u221a ,_ w V \u00b7 ,lm 9 T 4~0 ( + \u0398 \u03bb ] (12) The angle 9 can be seen in Fig.3 and Fig 4. It is important to notice that equations (10) through (12) have two solutions. This is because the rotors spin in different directions as can be seen in Fig. 4. Different rotation directions have the opposite effect on torques which produces the w sign in torque equations. These differences, induced from specific quadrotor construction, along with the augmented momentum equation function (2) provide an improved insight to quadrotor aerodynamics. Regardless of the flying state of the quadrotor, using these equations we can effectively model its behavior. III. HYBRID CONTROLLER OVERVIEW When developing a controller for this type of highly nonlinear system, two main problems arise", + " The dormant state, in which the flip controller waits for instructions from the user. Upon receiving a looping command from the superior controller the automaton moves to the second state. 2. The fast liftoff state triggers a request for maximal rotor spin. This causes the aircraft to liftoff with maximum climb speed. This state also switches off the height controller, causing the aircraft to fly in an open loop control. When vertical speed of the aircraft reaches 2.6m/s the controller switches to the next state. 3. The spin state turns off the rotors one and two (Fig. 4), thus making the quadrotor to spin around Y axis. This state is active for a certain amount of time +. During that time the aircraft rotation accelerates to a speed \u2126 . Assuming that the air resistance is much smaller than the momentum induced by rotors, we can easily calculate the time +. The next trigger arises upon the expiration of + and the controller switches to the fourth state. 4. The spinning state shuts down the propulsion system. When the aircraft makes the whole 360\u02da loop the sensors trigger the next state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001981_tia.2012.2226551-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001981_tia.2012.2226551-Figure7-1.png", + "caption": "Fig. 7. VSC common-mode circuit.", + "texts": [ + " 5 and line-to-ground voltages on each phase, where \u23a1 \u23a3 va vb vc \u23a4 \u23a6 = vb 2 \u00b7 \u23a1 \u23a3 ha hb hc \u23a4 \u23a6 (1) \u23a1 \u23a3 vu vv vw \u23a4 \u23a6 = vb 2 \u00b7 \u23a1 \u23a3 hu hv hw \u23a4 \u23a6 . (2) The switching function for each phase is a series of pulses that alternate between +1 and \u22121 at the switching frequency with pulsewidths over one switching period that vary according to the modulation functions for each desired phase voltage. The VSC-based VFD always has ha + hb + hc = 0 and hu + hv + hw = 0 because of the odd number of throws. As a result, the common-mode voltages, which are the sum of the three-phase voltages, are always nonzero. The common-mode circuit shown in Fig. 7 demonstrates that the common-mode current, and hence the emissions, can never be zero. The result is that EMI filters must always attenuate common-mode voltages which have significant frequency content at the switching frequency and harmonics of the switching frequency. Equations (1) and (2) demonstrate that pulsed voltages are always applied to the source and sink sides of the VSC. On the source side, high-frequency voltage content gets reflected onto the LISN, so EMI testing will deem the system incompatible unless capacitors are added across the lines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000076_j.robot.2008.01.002-Figure17-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000076_j.robot.2008.01.002-Figure17-1.png", + "caption": "Fig. 17. Ditch crossing when Rx /3 < d \u2264 Rx /2: (a) initial state, (b) swing leg 4, leg 2 and leg 3 and (c) move the body.", + "texts": [ + " 15, the optimal gait planning is that the quadruped robot uses the proposed twophase discontinuous gait for crossing ditches with 0 < d \u2264 Rx/3, and uses the previous one-phase discontinuous gait for crossing ditches with Rx/3 < d \u2264 Rx/2. If xo is long enough for the quadruped robot to lapse more than one stride length before confronting the ditch, the robot will have the two-phase gait in that segment for having fewest steps. In order to carry out this gait planning, gait transition procedures from the two-phase gait to the one-phase gait and stride length adjustment should be assured beforehand. For instance, consider Fig. 17(a) where the quadruped robot, having walked with the proposed two-phase gait with r1(\u2212Rx/6 \u2264 r1 < Rx/6), is about to cross a ditch with the width d(Rx/3 < d \u2264 Rx/2). More specifically, we assume that r1 is less than 0, i.e., \u2212Rx/6 \u2264 r1 < 0. Based on the proposed gait planning, the quadruped robot should change its gait to one-phase for getting ditch crossing ability. The stride length of the one-phase gait with \u2212Rx/6 \u2264 r1 < 0 is Rx/2 from (2). As Fig. 17(a) indicates, the quadruped robot should end the last cycle of the two-phase gait in such a way that the edge of the ditch is onto the boundary of the working areas of the front legs, that is, the distance between the foot point of leg 1 and the ditch is set to be Rx/2 + |r1|. Because the center of gravity is on the side of the support pattern made of r2 and r3 in the initial state, leg 4 can be moved onto the mirror point of leg 1. After leg 4 is placed on \u2212r1, leg 2 and leg 3 are moved on the front boundaries of their working areas one by one. The result of these procedures is illustrated in Fig. 17(b). The gait transition is accomplished in the final state of Fig. 17(c), where the center of gravity is translated onto C0 with the passive swing of leg 1. (r10, r20, r30, r40) indicates the set of four foot points after the translation of the robot body. By the leg sequence shown in Fig. 5, the quadruped robot will cross the ditch with the stride length \u03bb (=Rx/2) and its foot points will be moved onto (r11, r21, r31, r41). We can easily deduce from Fig. 17(c) that the maximum width of the ditch that the quadruped robot can cross is Rx/2 \u2212 |r1|, which corresponds to the triangular shape of the one-phase gait shown in Fig. 15. This paper has proposed an alternative fault-tolerant gait for a quadruped robot with a locked joint failure. The proposed gait is a kind of two-phase discontinuous gaits characterized by the sequential motion of legs and the robot body. For realizing fault tolerance against the locked joint failure, the leg and body movement sequence has been adapted such that it includes passive swing of the failed leg twice associated with the translation of the robot body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002500_j.cpc.2011.03.004-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002500_j.cpc.2011.03.004-Figure1-1.png", + "caption": "Fig. 1. Asymmetrically shaped micro-objects used in the simulations: a) arrowshaped shuttle; b) 8-teeth gear; c) 16-teeth gear.", + "texts": [ + " The center of mass of the gears is kept fixed at the center of simulation box, while only rotational movements around an axis perpendicular to the (x, y) plane are allowed. As a consequence the gear equation of motion is given by the second of Eqs. (1), which becomes a scalar equation for the z component of the angular velocity. The rotational mobility of the gear K\u22a5 is estimated from that of a thin disk of equal radius [9]. We investigate different shaped gears, varying the number and size of the teeth as well as the overall object shape \u2013 see Fig. 1 b) and c). Once immersed in the bacterial bath gears start to move, giving rise to a unidirectional rotational motion, characterized by a fluctuating angular velocity with non-zero mean. While from a thermodynamical point of view the rectification process can be explained in terms of out-of-equilibrium and broken symmetries mechanisms, from a microscopic dynamical point of view it can be viewed as an effect of pushing bacteria which preferentially get trapped in the concave corners of the object boundaries. The asymmetric boundary then provides the preferential locations where motile cells get stuck and push the object in a net unidirectional motion. Gears angular velocity depends on object shape and on bath properties (presence of tumbling, considering or not hydrodynamic interactions). Considering the first 8-teeth gear (Fig. 1 b)), with average radius R g = 19.5 \u03bcm, we observe an average angular ve- locity of about 0.36 rad/s. Switching off the hydrodynamic interaction term the velocity decreases to the value 0.21 rad/s. The hydrodynamic interactions are then responsible of a more efficient pushing mechanism, allowing a more efficient alignment of bacteria along object boundaries. Considering, instead, a bath of bacteria without tumbling, the angular velocity increases at the value 0.49 rad/s. Tumbling acts as a noise disturbing the particles motion: cells leave their pushing positions more frequently than in the case of absence of tumbling", + " It is worth noting that bacteria mainly point inward and have net anticlockwise tangent velocity (the radial propelling force, and then velocity, is balanced by the force exerted by the object boundaries). In Fig. 4 the average tangent velocity of bacteria is shown as a function of r for the two cases of 16- and 8-teeth gear. It is evident a persistent unidirectional microswimmers velocity also outside the gears, indicating a coherent motion of bacteria up to about 3 bacterial lengths. We now analyze the case of asymmetric shuttles free to move along their axis (the x-axis) without rotational motions. We investigate a shaped shuttle with a concave back side (see Fig. 1 a)) and an arrow shaped front side \u2013 linear dimensions Lx = 15 \u03bcm and L y = 18 \u03bcm. The equation of motion is now given by the first of Eqs. (1), which becomes a scalar equation for the x-component of the shuttle velocity. The translational mobility of the shuttle Ms is roughly estimated from that of a thin disk [9]. Switching on the simulation, we observe again a unidirectional linear motion of the object, with a fluctuating velocity having a non-zero average value of 3.3 \u03bcm/s (see Fig. 2, right part)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001843_1.39537-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001843_1.39537-Figure2-1.png", + "caption": "Fig. 2 Thruster layout [10].", + "texts": [ + " (1\u20133) and (9) can be written as x f x; _x Bu (10) where f 2 _ _y y 2x x=kxi _ y < j\u0302 zk\u0302k3 2 _ _x x _ 2 y y <=kxi _ y < j _ zk\u0302k3 <=k> |X\u0307|, then the lateral component of the friction force is Tx \u2248 \u2212T0 V X\u0307, where T 0 is the break-out force. The lateral velocity X\u0307 makes the friction force T 0 to slightly deviate from the direction along the belt in proportion to X\u0307/V " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003547_1.4005013-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003547_1.4005013-Figure1-1.png", + "caption": "Fig. 1 Rotor with a ball balancer", + "texts": [ + " Even after such a long history of research, the following three fundamental problems have yet to be solved. (1) What is the theoretical response curve of a rotor with a two-ball balancer? (2) What kind of friction influences the accuracy of balancing? (3) How can we eliminate the effect of friction and attain sufficient balancing? The aim of the present paper is to clarify these three questions. In addition, we show a ball balancer which overcomes the abovementioned malfunctions of conventional ball balancers. 2.1 Theoretical Model and Equations of Motion. Figure 1 shows the theoretical model. A disc of mass M is mounted at the middle of an elastic shaft with spring constant k. The rotational speed x of the shaft is assumed to be constant. Let the geometrical 1Corresponding author. Contributed by the Technical Committee on Vibration and Sound of ASME for publication in the JOURNAL OF VIBRATION AND ACOUSTICS. Manuscript received December 6, 2009; final manuscript received June 29, 2011; published online January 18, 2012. Assoc. Editor: Cheng-Kuo Sung. Journal of Vibration and Acoustics APRIL 2012, Vol" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000152_0022-4898(81)90016-1-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000152_0022-4898(81)90016-1-Figure10-1.png", + "caption": "FIG. 10. Forces on front wheels of a two-wheel drive tractor when turning.", + "texts": [ + " I T Y N U M B E R - B M o B (d) Coefficient of roiling resistance at 10 \u00b0 slip angle against the four different of the side force-slip angle curve, and the rolling resistance, at low values of slip angle. The rolling resistance for these calculations can be taken to be that at zero slip angle which can already be satisfactorily predicted. For side-slope stability the forces at higher slip angles may be required (e.g. the stability of a two-wheel drive tractor \"crabbing\" across a side-slope with its front wheels pointing up the slope). Forces at higher slip angles will certainly be required for handling calculations. Figure 10 shows the forces acting on the front wheels of a two-wheel drive tractor when turning. The tractor is assumed to have a steering linkage which can provide perfect steering, i.e. a line drawn through the rear axle meets lines drawn perpendicular 42 D. GEE-CLOUGH and M. S. SOMMER STEERING FORCES ON UNDRIVEN, ANGLED WHEELS 43 J 2 0 IO I d & 4.00 \u2022 8 TYRE CONE INDEX VALUE LOAD SYMBOL (kPo) ( N ) EMoB IDMOB x 300 I 0 0 0 5 0 3 0 .035 o 150 150 8 . 4 6 0 . 0 4 6 ~, 5 5 0 I o 0 0 9 . 6 6 0 . 0 6 6 D 9 0 0 I 0 0 0 14" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure13-1.png", + "caption": "Figure 13: Mechanical Hybrid Variator Unit", + "texts": [], + "surrounding_texts": [ + "Traditional Turbo Compounding systems comprise a second power turbine located in the exhaust downstream of the primary turbine. This second turbine is mechanically connected to the crankshaft of the engine in order to recover energy from the exhaust that would have otherwise have been lost. Turbo compounding has been applied to airplane engines and in both on- & off-highway Commercial Vehicles (e.g. Case Steiger and Detroit Diesel DD15). An alternative arrangement to introducing a second power turbine is to introduce a mechanical variable drive system connected between the primary turbine and the crankshaft of the engine (Figure 10). By extending a shaft from the primary turbine, a Torotrak Variator can be connected to the shaft and subsequently connected into the crankshaft. Gearing either side of the Variator, together with the spread of the Variator itself, provide the necessary ratio to connect the low speed crankshaft to the high speed turbocharger. The advantage of utilising a Torotrak Variator in this system is the ability to control both the direction of torque (and hence energy) flow as well as the quantity of energy. If reaction force is applied on the pistons of the Variator in one direction, energy will flow from the turbo charger into the crank i.e. energy can be recovered from the exhaust gas to the crankshaft. Conversely, if reaction force is applied to the pistons in the opposite direction, then energy will flow from the crankshaft into the turbocharger i.e. the crank will drive the turbo and hence supercharge the engine. Therefore, by modulating the loads on the Variator, the turbocharger can : 1. Operate as a supercharger at low exhaust gas flows. 2. Operate as a standard turbocharger at \u201cmedium\u201d exhaust gas flow with the Variator unloaded. 3. Load the turbocharger to recover energy to the crank at high exhaust energy gas flows. 4. Enhance engine braking by increasing cylinder pressure (supercharging) in overrun. A Turbo Compounding system of the type described above was developed by Torotrak a number of years ago for a Commercial Vehicle application. This system which successfully met the project targets of : 1. Improving the thermal efficiency of the engine. 2. Supercharge the engine at low engine speeds. 3. Recovering energy to the crankshaft at high exhaust gas flows. A twin cavity Variator utilising three rollers per cavity was developed and successfully applied to the target engine. Hardware is shown in Figure 11. To damp out the torsional oscillations from the crankshaft, a fluid coupling was utilised between the crank and the Variator. Although successful in meeting the project targets, the Torotrak technology was immature at the time of this project (early 1990\u2019s). The Variator utilised a twin cavity, six roller Variator design of 85mm roller diameter for a ~ 45kW capacity. The result, therefore, was a variable drive system that was too big, heavy and expensive for application as a Series Production Turbo Compounder Drive. Since this time, Torotrak have developed the full toroidal traction drive technology with particular focus on cost, weight, package and performance [1],[2], [5] & [6]. An example of the level of development is demonstrated in the Flywheel Based Mechanical Hybrid system developed for Formula 1 and mainstream automotive applications. The Flywheel Based Mechanical Hybrid system comprises a flywheel as the energy storage medium and a Torotrak full toroidal Variator to control the power and energy flow between the vehicle and flywheel & is described separately [6]. (Reproduced courtesy of Flybrid Systems LLP) The result of the developments is a particularly power dense Variator unit with the 55mm roller diameter unit from the Mechanical Hybrid having a maximum power capacity in excess of 110 kW. The hardware from Figures 12 and 13 has been realised and the Variator power capacity, efficiency, durability, contact pressure and temperature performances have been measured \u2013 see Figures 14 & 15. From this data, it is clear that the full toroidal traction drive Variator technology has progressed to a level where it is suitable for series application as a variable drive for a turbo compounding system. Analysing the power capacity of two Variator designs with 50mm or 60mm roller diameter Variators produces Figures 16 & 17 i.e. Variators capable of 100kW and 130kW respectively with efficiencies of 90 to 91%." + ] + }, + { + "image_filename": "designv11_3_0003237_s11665-013-0583-2-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003237_s11665-013-0583-2-Figure10-1.png", + "caption": "Fig. 10 Schematic representations of the columnar-grain microstructure in: (a) untransformed austenite; (b) austenite partially transformed into allotriomorphic ferrite; and (c) austenite partially transformed into allotriomorphic and Widmanstatten ferrite", + "texts": [ + " In the remainder of this section, procedures are presented for computing the final volume fractions of the phases formed during austenite decomposition within different portions of the FZ. Allotriomorphic-Ferrite Volume Fraction. Austenite grains formed during solidification are, for the most part, of a columnar shape and aligned with the direction of the maximum heat extraction. For modeling purposes, the cross section of the columnar austenite grains is typically idealized as being of a regular hexagonal shape. In other words, as shown schematically in Fig. 10(a), the columnar austenite grains are assumed to form a perfect honeycomb structure, with the axes of the hexagonal prisms being aligned with the local maximum heatextraction direction. In addition to specifying the shape of the austenite grains, the hexagonal-section edge length, a, must be specified since, as will be shown below, it affects the final volume fraction of the allotriomorphic ferrite. The hexagonalsection edge length is, in turn, mainly affected by the maximum local cooling rate attained by the liquid during solidification", + " On the other hand, at sufficiently low temperatures, the diffusivity of carbon becomes quite low, so that transformation of austenite fi allotriomorphic ferrite becomes kinetically constrained, causing a reduction in a1; and (d) finally, it is assumed that a functional relationship can be established between the allotriomorphic ferrite plate thickness, qf, and its volume fraction, VaA. This functional relationship is inferred by applying a simple geometrical computational procedure to the schematic displayed in Fig. 10(b) which depicts a partially transformed section of austenite. The resulting functional relationship can be stated as: VaA \u00bc 2qf tan\u00f030 \u00de 2a 2qf tan\u00f030 \u00de\u00f0 \u00de\u00bd a2 : \u00f0Eq 6\u00de Widmanstatten-Ferrite Volume Fraction. Examination of Fig. 8 reveals that, at relatively high transformation temperatures, the onset of austenite to Widmanstatten ferrite phase transformation causes the kinetically sluggish austenite to allotriomorphic ferrite transformation to cease. Following a detailed analysis presented in our recent work (Ref 5), the growth rate of the Widmanstatten ferrite is assumed to be controlled by the rate of lengthening of this lens-shaped phase in a direction normal to the local allotriomorphic ferrite/austenite interface. To help clarify geometrical/topological details related to the formation of Widmanstatten-ferrite, a simple schematic of partially transformed austenite grains is depicted in Fig. 10(c). Examination of this figure reveals the presence of prior austenite grain-boundary regions which have been transformed into allotriomorphic ferrite, as well as lenticular-shaped Widmanstatten plates advancing from the allotriomorphic ferrite/austenite interfaces toward the untransformed austenite grain centers. As discussed in great detail in our recent work (Ref 5), the rate of the austenite fi Widmanstatten-ferrite phase transformation is affected not only by the para-equilibrium condition still present at the ferrite advancing front and the associated carbon diffusion from this front into the untransformed austenite, but also by the displacive character of the austenite to Widmanstatten ferrite transformation. Following the procedure described in Ref 10, 48, which is based on the calculation of the Widmanstatten-ferrite area fraction within the austenite grains with hexagonal crosssection, Fig. 10(c), the following expression is derived for computing the Widmanstatten-ferrite volume fraction, VaW: VaW \u00bc C4G 2a 4qf tan\u00f030 \u00de t2aW \u00f02a\u00de2 \" # ; \u00f0Eq 7\u00de where C4 [= 7.367 s 1 (Ref 19)] is an alloy-composition independent constant, G [= 52 lm/s (Ref 47)] is the Widmanstatten Journal of Materials Engineering and Performance ferrite lengthening rate, and taW is the total time available for the austenite fi Widmanstatten ferrite transformation (it should be recalled that once temperature drops below BS, the austenite fi Widmanstatten ferrite transformation ceases and it is replaced with a austenite fi bainite phase transformation)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure9-1.png", + "caption": "Fig. 9 By using an idler loop the 2DOF parallel manipulator can be balanced with two CRCMs near the base", + "texts": [ + " This means that there are only two CRCMs necessary, which oth can be constructed compactly near the base, as shown in Fig. . This is a configuration described in Ref. 14 . The former RCMs become fixed countermasses. It is also possible to derive this parallel mechanism by combiation of an idler loop 15 and a CRCM-balanced double penduum, as shown in Fig. 8. Also in this case, only two CRCMs are ecessary, and by using the countermass of the idler loop as a RCM, they can be constructed near the base. This result is hown in Fig. 9 and it has only one fixed countermass instead of wo, as in the configuration of Fig. 7. With the equations of the angular momentum of the balanced ouble pendulum being known, the inertias of the CRCMs and the nertia equations of the mechanism can be calculated quickly by imply adding the equations of each individual double pendulum. s an example, the angular momentum and inertia equations of A O ig. 6 2DOF balanced parallel mechanism obtained by combiation of two CRCM-balanced double pendula he parallel mechanism of Fig", + "org/ on 01/28/201 dulum, shown with chain driven CRCMs in Fig. 14. Also for four-bar mechanisms, the different CRCM configurations are applicable just as the substitution of the well-known kinematic relations into the inertia equations of the double and single pendula to obtain the inertia about one of the links. This means that with the equations for the double and single pendula and the kinematic relations the inertia of any four-bar mechanism can be written down easily. It is a special case for which the four-bar linkage becomes a parallelogram. From Fig. 9, and assuming the link between O and A to be fixed with the base, the resulting parallelogram can be balanced, as in Fig. 15, with solely a CRCM. Because the coupler link does not rotate, its center of mass can be located arbitrarily. 5 3DOF Parallel Mechanisms CRCM-balanced 3DOF planar and spatial mechanisms can be synthesized by combining the CRCM-balanced double pendula. Two examples are the planar 3-RRR parallel mechanism of Fig. 16, which has one rotational and two translational DOFs, and the spatial 3-RRR parallel mechanism of Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002588_1.4005215-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002588_1.4005215-Figure6-1.png", + "caption": "Fig. 6 Details of cooling channels on the rotor. (a) Leading edge region. (b) Cross section of simplified model of the leading edge region.", + "texts": [ + " The mass balance equation for the high pressure side is _mcooling \u00fe X _mleakage H \u00bc X _mch bump \u00fe X _mch rotor \u00fe X _mto LE groove (6) where _mch bump is the mass flow rate through the bump cooling channels, _mch rotor is mass flow rate through the rotor cooling channels, _mto LE groove is the mass flow rate through the leading edge grooves, and _mleakage H is a leakage flow from the bearing to the high pressure plenum. The cooling air channels in the rotor are circumferentially arranged multiple (Nch rotor) axial channels as shown in Fig. 6. In addition, _mcooling in Eq. (6) is the cooling air flow rate supplied to the high pressure plenum through multiple (Ncooling) ducts formed on the housing _mcooling \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi N2 cooling 2 fcooling dcoolingA2 coolingqcooling Pcooling PplenumH Lcooling s (7) where fcooling \u00bc 0:316=Re0:25 or 64=Re depending on the flow condition (flow Reynolds number through the main cooling air port is found in a iterative way while satisfying the mass continu- ity at the plenum), Acooling is a cross section area of the cooling air supply hole" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003469_j.oceaneng.2013.01.001-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003469_j.oceaneng.2013.01.001-Figure8-1.png", + "caption": "Fig. 8. Simplified model of 4-legged robot system.", + "texts": [], + "surrounding_texts": [ + "In this section, the final goal is to establish the unified mathematical framework for analysis of the mobility and agility of the legged robot from given torque bounds, frictional constraints, and hydrodynamic forces. As shown in Fig. 5, the procedure of mathematical framework consists of torque constraint equation, linear program, and dynamic equation. Given joint torque bounds are limited by the torque constraint Eq. (37) and the limited maximum torque bound described as a polygon with vertices is extracted by conventional LP (Linear Programming). The inside torques of the polygon satisfy no slip condition at contact point. The vertices of the polygon are applied to (28), then, we can get a polytope formed by the vertices of admissible acceleration of the body. The linear components of the acceleration are related to mobility, its rational components are related to agility." + ] + }, + { + "image_filename": "designv11_3_0001007_0020-7403(75)90014-4-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001007_0020-7403(75)90014-4-Figure1-1.png", + "caption": "FIG. 1. Schematic representation of vehicle.", + "texts": [ + ") The paper is organized as follows: section 2 contains a description of the vehicle and the definitions of symbols used in the analysis that follows. In section 3, five relationships associated with the fact that both wheels must be in contact with the supporting sur face are derived. Five further fundamental equations are developed in section 4 by taking rolling without slip into account. Section 5 deals briefly with steady turning. Finally, a comprehensive discussion of re sults is presented in section 6. Fig. 1 is a schematic representation of a single track vehicle consisting of a frame A, fork B, rear wheel 0 and front wheel D. The vehicle is des cribed geometrically by the constants a, b, r, r' and (), the rake angle. The \"configuration\" of the vehicle is charac terized by the steering angle a between the central planes of A and B, and by the wheel rota tion angles y and o. For the configuration depicted in Fig. 1, a is equal to zero while y and 0 are positive. The orientation of A in a \"fixed\" reference frame N depends on three angles generated as follows: Let a1' 1102 and a 3 = a 1 x 1102 form a set of orthonormal unit vectors fixed in A as shown in Fig. 1, and let n1 , n 2 and n3 = n1 x n2 form a similar set of unit vectors fixed in N, with n1 normal to the plane that supports the vehicle. Next, align a i with n i (i = 1,2,3) and then subject A successively to an a 1 rotation of amount K, an a 2 rotation of amount A, and an a3 rotation of amount iL' per forming each of these rotations in the right-handed sense about an axis parallel to the unit vector under consideration in its current orientation. The three angles are called the yaw angle (K), the roll angle (A) and the pitch angle (iL)' They are shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001282_aim.2009.5230018-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001282_aim.2009.5230018-Figure1-1.png", + "caption": "Fig. 1. Twin rotor multi-input multi-output system.", + "texts": [ + " Based on the simulation results, the new approach improves the performances in setpoint and tracking control a lot. This paper is organized as follows. Section II presents the twin rotor multi-input multi-output system model and brief introduction of the TRMS system. Section III comprehensively discusses the method of designing the hybrid PID controller. In section IV, both the simulation in decoupled condition and simulation in cross-coupled condition of the TRMS are presented, followed by some concluding remarks in section V. The TRMS, as shown in Fig. 1, is characterized by complex, highly nonlinear and inaccessibility of some states and outputs for measurements, and hence can be considered as a challenging engineering problem [12]. The control objective is to make the beam of the TRMS move quickly and accurately to the desired attitudes, both the pitch angle and the azimuth angle. The TRMS is a laboratory set-up for control experiment and is driven by two DC motors. Its two propellers are perpendicular to each other and joined by a beam pivoted on its base that can rotate freely on the horizontal plane and vertical plane" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001713_s11431-009-0064-x-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001713_s11431-009-0064-x-Figure10-1.png", + "caption": "Figure 10 The 3D profile of a fish body with a caudal fin.", + "texts": [ + " z x x x z x x x z x x x x x = \u00b1 +\u23a7 \u23aa \u2212\u23aa \u23aa = \u00b1 \u2212 \u2212\u23aa \u23a8 < = \u00b1 \u2212 + <\u23a9 \u2264 \u2264 \u2264 \u2264 \u23aa \u23aa \u23aa \u23aa (28) As shown in Figure 9, the leading edge and trailing edge profiles are defined by 3 2 l l LE l l l l 3 2 l l TE l l l ( ) 39.543 3.685( ) 0.636 0.7, 0.15 0.15; ( ) 40.74 9.666( ) 0.77, 0.15 0.15. z x z z z z z x z z z \u23a7 = \u2212 + + \u23aa \u2212\u23aa \u23a8 = \u2212 + +\u23aa \u23aa \u2212\u23a9 \u2264 \u2264 \u2264 \u2264 (29) In chord-wise sections, the shape of a caudal fin takes the shape of NACA 0040 instead of NACA 0016 in ref. [11]. The 3D profile of a fish body with a caudal fin is shown in Figure 10. 664 WU ChuiJie et al. Sci China Ser E-Tech Sci | Mar. 2009 | vol. 52 | no. 3 | 658-669 3.3.2 Flapping rules. It is assumed that bending of the fish body happens only within the xl-yl plane, and the tail is assumed to oscillate and rotate with a given or computed rule[11]. In the self-propelled swimming, the motion of fish body is fully described by specifying the motion of its backbone with several key kinematic parameters. The motion is characterized by a traveling wave with a smoothly varied amplitude" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure5-1.png", + "caption": "Fig. 5 Principle of the setup of a grinding wheel with an internal cooling lubricant supply", + "texts": [ + " As a compromise between good coolant supply and feasible undercuts, a setting angle of 45 degree has been specified. To guarantee the machinability of the undercuts, a minimal diameter of the milling tool of 3 mm is required. The resulting width of the coolant outlet is approximately 4 mm. Finally, the cross-sectional areas of the outlets and in the transition areas were held constant to assure constant flow velocities. Grinding wheels with an internal cooling lubricant supply must consist of a base plate and a cover plate to cover the channels (see Fig. 5). In this case, the channels and the outlets of the channels are restricted to the base plate of the grinding wheel. This means there are no channels inside the cover plate. The channels inside the base plate decrease the stability of the grinding wheel. For this reason, the width of the base plate is set to 15 mm, which assures sufficient stability and stiffness in high speed grinding (this was checked by stability calculations made by a grinding wheel manufacturer). The height of each coolant outlet is 14" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002012_j.triboint.2010.06.002-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002012_j.triboint.2010.06.002-Figure1-1.png", + "caption": "Fig. 1. Squeeze film geometry: enlarged pictures in circles indicate the unidirectional roughness pattern in y-direction (longitudinal roughness).", + "texts": [ + " It is shown that the applied roughness and magnetic field enhances the pressure, Nomenclature a length of the plate b width of the plate B0 applied magnetic field in z-direction d a/b aspect ratio h Non-dimensional film thickness h0 initial film thickness c \u00bc c h0 non-dimensional roughness parameter d \u00bc d a Thickness of the porous facing k permeability of porous facing mu porosity M0 Hartmann number, B0h0ffiffiffiffiffiffi s=m p E(p), P pressure in the film and porous region, respectively p dimensionless pressure\u00bc E\u00f0p\u00deh3 ma2dh=dt t time u,v,w velocity components in film region u,v, w velocity components in porous region W E\u00f0w\u00deh3 ma3bdh=dt , Dimensionless load x,y,z rectangular co-ordinates am mp a bn np b gmn p a \u00f0m 2\u00fed2n2\u00de 1=2 gmn agmn m viscosity of fluid s conductivity of fluid c0 kd h3 0 , permeability parameter DT h2 0 E\u00f0w\u00de ma3b Dt dimensionless squeeze film time load carrying capacity and time of approach and the results are compared with corresponding non-magnetic and non-roughness case [14,16]. 2. Mathematical formulation and solution of the problem A schematic diagram of the squeeze film geometry is shown in Fig. 1. Consider a squeezing flow between two rectangular plates when the upper plate has a porous facing is assumed to move while the lower plate remains fixed. A uniform magnetic field B0 is applied perpendicular to the two plates with the account of velocity slip at the porous facing [17]. The flow in the porous media follow the modified Darcy\u2019s law due to Horia [18], while in the film region the usual assumptions of hydromagnetic lubrication theory [19] hold good. Following the assumptions of Wu [14] and the usual assumptions of magneto hydrodynamic lubrication [19], the basic equations governing the hydromagnetic law of the lubricant in different regions are: For the film region: @2u @z2 M2 0 h2 0 u\u00bc 1 m @p @x \u00f01\u00de @2v @z2 M2 0 h2 0 v\u00bc 1 m @p @y \u00f02\u00de @p @z \u00bc 0 \u00f03\u00de @u @x \u00fe @v @y \u00fe @w @z \u00bc 0 \u00f04\u00de For the porous region: u\u00bc k m @P @x 1 1\u00fe k mu M2 0 h2 0 \u00f05\u00de v\u00bc k m @P @y 1 1\u00fe k mu M2 0 h2 0 \u00f06\u00de w\u00bc k m @P @z \u00f07\u00de @u @x \u00fe @v @y \u00fe @w @z \u00bc 0 \u00f08\u00de The relevant boundary conditions for the velocity components are u\u00bc v\u00bc 0 at z\u00bc 0 \u00f09\u00de u\u00bc v\u00bc 0 at z\u00bcH \u00f010\u00de The solution of Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002485_1.4025350-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002485_1.4025350-Figure1-1.png", + "caption": "Fig. 1 Mechanical components of a planetary needle bearing setup", + "texts": [ + " The most common type of planet bearings used in automotive planetary gear sets are needle bearings, while complete cylindrical roller bearings (with the outer and inner race as a unit) are used for larger-scale planetary gearboxes used in applications such as helicopter transmissions and wind turbine gearboxes. A planet needle bearing consists of a set of needles either held by a cage or not (in the case of full complement needle bearings) with the planet pin and the planet bore acting as the inner and outer races, respectively, as illustrated in Fig. 1. Full complement bearings are used in higher load applications. Caged needle bearings are used for higher speed applications and will be the focus of this study. While the intended application of this model is prediction of power losses of planetary gear set planet bearings, the proposed model can be used for any caged cylindrical roller bearing under arbitrary radial force and moment loading conditions. As in any geared transmission, power losses of a planetary gear set can be broken into two main categories, load-dependent (mechanical) and load-independent (spin) power losses as P \u00bc PLD \u00fe PLI", + " Manuscript received June 3, 2013; final manuscript received August 21, 2013; published online September 20, 2013. Assoc. Editor: Qi Fan. Journal of Mechanical Design DECEMBER 2013, Vol. 135 / 121007-1Copyright VC 2013 by ASME Downloaded From: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jmdedb/927630/ on 02/01/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use and load distributions. When used with helical gears, planet bearings shown in Fig. 1 are subjected to not only a radial force but also overturning moments. Literature lacks combined load distribution and mechanical power loss models for cylindrical roller or needle bearings under such loading conditions. The influence of other needle bearing features such as needle crowning on rolling power losses has also not been investigated in these previous studies. The proposed research aims at developing a mechanical power loss model for cylindrical roller or needle type caged planet bearings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001139_09511921003667698-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001139_09511921003667698-Figure5-1.png", + "caption": "Figure 5. Different types of distortions of a rectangular welded plate (not to scale).", + "texts": [ + "760 42 20 30.0 80 0.65 7 2.456 43 14 30.0 80 0.35 7 2.456 44 14 39.0 80 0.35 11 2.456 45 14 30.0 182 0.65 7 2.456 46 17 34.6 130 0.50 11 3.760 47 20 39.0 182 0.35 7 5.635 48 17 34.6 130 0.50 9 3.760 49 20 39.0 80 0.65 11 2.456 50 20 39.0 80 0.35 11 5.635 51 14 39.0 182 0.35 7 2.456 52 17 34.6 130 0.50 9 3.760 53 20 34.6 130 0.50 9 3.760 D ow nl oa de d by [ N or th ea st er n U ni ve rs ity ] at 1 2: 22 2 4 N ov em be r 20 14 distortion; 4) bending distortion. These distortions are schematically shown in Figure 5. Note that the longitudinal shrinkage and bending distortion are generally small. Therefore, in this research, transverse shrinkage and angular distortion are considered for the study. The plate layout for distortion measurement is shown in Figure 6. Two plates were tack welded at the two ends, as shown by points A and B in Figure 6, to measure the transverse shrinkage and angular distortion. Transverse shrinkage was measured at L1, L2 and L3 locations (Figure 6) by using a vernier calliper (111\u2013 332; Mitutoyo Corporation, Kanagawa, Japan) and their average was calculated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000796_rd.186.0534-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000796_rd.186.0534-Figure2-1.png", + "caption": "Figure 2 Taper-flat steel slider used in the experimental study.", + "texts": [ + " Prior to the experiment, the three pads of the slider were lapped flat to within a 50 nm peak-to-valley sur face roughness across the flat portion, leaving small tapered sections at the leading edges. The steady-state flying characteristic of the slider was then determined on a glass disk using white light interferometry. The glass disk used for the spacing calibration was a pre cision finished disk with less than 25 nm peak-to-valley surface roughness. Since the slider-disk separation was 535 for this investigation and shown in Fig. 2. The slider load, applied by lowering the slider support against the disk, is measured by means of strain gauges mounted on the suspension spring. The rotating disk is maintained at solid ground poten tial through a connection in a mercury bath mounted on the axis of the spindle shaft. Adequate grounding was found to be essential for eliminating electrical noise due to charge buildup from the driving belt. Contacts be tween the slider and the disk are detected and counted using the circuit shown in Fig", + " We observe that at low veloci ties the measured spacings depart from the numerically predicted ones, seemingly because of material interac tions due to surface roughness and increased inaccuracy of the theoretical model which assumes perfectly flat surfaces. After the steady state flying characteristics was established, the glass disk was replaced by a metallic disk, thus permitting detection of slider/disk contacts and the investigation of the nature of the transition from continuous sliding to steady flying. A typical set of voltage output traces is shown in Fig. 5 as a function of velocity, using the slider depicted in Fig. 2. It can be seen clearly from these photographs that contacts, corresponding to narrow spikes, are present over most of the examined speed range. These contacts decrease in number as well as in magnitude as the veloc ity increases, and disappear completely above approxi mately 2000 cm/ s. A similar behavior is also observed if the load of the slider is increased, although in this case the absence of contacts is observed at a much higher disk velocity. Characteristic of all the observed traces is the random nature of the pulses and the weak dependence of the pulse width on the sliding speed", + " From this viewpoint, then, the results of the electrical resistance method should be taken as addi tional detailed information that is not obtainable by other methods. It is apparent that continued contact between the sli der and the disk results in wear of the contacting sur faces. Therefore, additional information concerning the transition to fully hydrodynamic lubrication may be ob tained by measuring the amount of wear of the slider and the disk. In Fig. 7 the volume V, worn off the trailing pad of the slider shown in Fig. 2, is plotted as a function of velocity for a constant sliding distance s. A particu late disk with a polymeric overcoat, containing finely dispersed alumina particles as the primary abrasive ma terial, was used in this study because it was found that wear, seemingly abrasive in nature, occurs for the parti- R. C. TSENG AND F. E. TALKE IBM J RES. DEVELOP. (d) culate coating/ steel slider interface mainly in the form of slider wear. In particular, slider wear was seen to consist of a gradual, uniform removal of material from the flat portion of the bearing surfaces [7], thus making it possible to calculate the total amount of wear by mea suring the increase in flat lengths, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003953_15325008.2014.880968-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003953_15325008.2014.880968-Figure4-1.png", + "caption": "FIGURE 4. Definition of rotor initial position.", + "texts": [ + " When the B phase torque main winding Nb and suspension windings Nx and Ny are energized with currents imb, ix, and iy, respectively, the magnetic equivalent circuit is shown in Figure 3, where Pb1\u2013Pb8 are the permeances D ow nl oa de d by [ U ni ve rs ity o f St el le nb os ch ] at 0 3: 26 0 5 N ov em be r 20 14 of the air gap between the stator and rotor tooth poles, \u03c6b1\u2013\u03c6b8 are the magnetic fluxes, and subscripts 1, 2, . . . , 8 denote the reference number of the stator poles shown in Figure 1. The mutual inductances between the three torque main windings can be ignored [17]; Pb9, Pb10, Pb11, and Pb12 are assumed as zero. The rotor rotates clockwise from the initial position \u03b8 = 0\u25e6 defined in Figure 4. The rotor position range is from 0\u25e6 to 15\u25e6 for the B phase torque main winding being energized. From Figure 3, one obtains\u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 \u03c6b1 Pb1 + Nmimb + Nxix = \u03c6b2 Pb2 \u2212 Nmimb + Nyiy = \u03c6b3 Pb3 + Nmimb \u2212 Nxix = \u03c6b4 Pb4 \u2212 Nmimb \u2212 Nyiy, \u03c6b5 Pb5 + Nxix = \u03c6b6 Pb6 + Nyiy = \u03c6b7 Pb7 \u2212 Nxix = \u03c6b8 Pb8 \u2212 Nyiy = \u03c6b1 Pb1 + Nmimb + Nxix . (1) \u03c6b1 + \u03c6b2 + \u03c6b3 + \u03c6b4 + \u03c6b5 + \u03c6b6 + \u03c6b7 + \u03c6b8 = 0. (2) The flux linkages \u03d5mb, \u03d5x, and \u03d5y, corresponding to the motor winding currents imb, ix, and iy, can be expressed as\u23a7\u23a8 \u23a9 \u03d5mb = (\u2212\u03c6b1 + \u03c6b2 \u2212 \u03c6b3 + \u03c6b4)Nm \u03d5x = (\u2212\u03c6b1 + \u03c6b3)Nx + (\u2212\u03c6b5 + \u03c6b7)Nx \u03d5y = (\u2212\u03c6b2 + \u03c6b4)Ny + (\u2212\u03c6b6 + \u03c6b8)Ny " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003558_powereng.2011.6036561-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003558_powereng.2011.6036561-Figure2-1.png", + "caption": "Fig. 2 Permanent Magnets Arrangement", + "texts": [ + "(n mag WindingsofNumber j jjflux fluxii \u03d5\u03c1\u03b8\u03d5\u03d5\u03d5 \u03d5\u03d5\u03d5\u03bb \u03c0 1 2 0 (11) Now, it is possible to derive an electrical system of equations for a PM machine: dt d)t(i.r)t(v \u03bb+= (12) Where, v, r, i and \u03bb are terminal voltages, windings resistance, windings current and flux linkage respectively. The parameters in (12) are defined by: ])(n.....)(n.[)](g[ ].....[ r00 0r0 00r r ]iii[i],vvv[v j1 T flux j1 T c b a cba T cba T \u03d5\u03d5\u03d5 \u03bb\u03bb\u03bb == \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 = == (13) The functions for the windings, magnet and permeance are calculated once at the beginning; during the simulation only the magnet function is properly shifted to reflect the rotation of the machine. Fig.2 shows the magnet arrangement in a PM machine at a particular\u03b8 respect to the stator Ref. The magnetic field intensity of the magnet (Hm) and its thickness (lm) are assumed to be constant thus, its MMF is evenly distributed in across and the relevant function can be determined theoretically by: \u23a9 \u23a8 \u23a7 = = wiseOther areamagnetizedInl.H),(MMF mmmag 0 \u03b8\u03d5 (14) On the other hand, according to (7) and (6), if the magnets are assumed to be the MMF source, then the magnet function is defined by: \u03d5 \u03d5 \u03b8\u03d5 \u03d5\u03c0 \u03b8\u03d5\u03b8\u03d5 \u03c0 \u222b\u2212 \u2212= 2 0 12 1 d )(g ),(MMF )(g ),(MMF),(F mag magmag (15) This equation emphasizes that the air gap geometry (especially the stator saliency) directly affects the magnet function by changing the F(0) value" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002789_cjme.2013.03.532-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002789_cjme.2013.03.532-Figure1-1.png", + "caption": "Fig. 1. Multiple pinions driving in TBM", + "texts": [ + " ZHOU, et al[4], studied the effect of injecting slurry inside a jacking pipe tunnel in silt * Corresponding author. E-mail: weijing@dlut.edu.cn This project is supported by National Basic Research Program of China (973 Program, Grant No. 2013CB035402) \u00a9 Chinese Mechanical Engineering Society and Springer-Verlag Berlin Heidelberg 2013 stratum using an experimental method. MARSHALL, et al[5], studied the influence of the moment and force caused by propulsion system. To bear huge loading, TBM usually uses multiple pinions driving to support the tunneling excavation (see Fig. 1). However, it can also cause driving torque distribution imbalance. Meanwhile, due to complex and varied geological conditions, TBM often bears the radial force, axial force and overturning moment simultaneously. During tunneling, the slewing bearing mainly suffers the external load of the cutter head and produces elastic deformation. For the driven gear and the ring gear of the slewing bearing, the deformation will lead to displacement disturbance because of the multi-point meshing. The accident of blocking may occur easily in the process of tunneling resulting from complicated geological conditions, and even fracture phenomena in small gear shafts and gear teeth will take place[6]", + " The load-sharing coefficient of multiple pinions driving is about 1.2 when cutter head rotational speed ranges from 0.5 rmin to 1.5 rmin, that is, despite of the cutter head rotational speed is very low, the dynamics meshing force and dynamic torque of each pinion are of oscillation trend around its static torque, with its amplitude being about 1.2 times of static torque. The load-sharing coefficients are not same when multiple pinions driving are not synchronized, which imply that torques in the four pinions, as shown in Fig. 1, are different. Supposing T is the average torque of all pinions at a certain time, the output torques in 1\u20134 driving pinions are, respectively, T1T, T21.4T, T31.4T, T40.6T and the total output torque is 4T. This is related to the case of rated cutter head rotational speed 0.9 rmin. The load-sharing coefficients of multiple pinions driving can be obtained using Runge-Kutta integration method based on equivalent mathematic model (Fig. 6). It can be seen from Fig. 6 that the load-sharing coefficient of each pinion is of oscillation trend around corresponding static torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001149_13552541011011668-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001149_13552541011011668-Figure3-1.png", + "caption": "Figure 3", + "texts": [ + " Therefore, the transmitted energy to the workpiece is a piece of valuable information to study when considering bond formation during UC. With that in mind, in the current study, an energy-based analytical model was developed to characterize the influence of process parameters on the transmitted energy during UC, and further to investigate any correlation between the transmitted energy and LWD. The objective of the current study is to characterize the energy input at the bonding interfaces during metal foil deposition on a base plate with UC. In UC, as shown in Figure 3, a metal foil is fed between the sonotrode and the base plate. The metal foil is pressed against the base plate by the sonotrode at a pre-set normal force. The sonotrode travels along the foil length direction at a pre-set traveling speed and oscillates perpendicular to the foil length at a pre-set amplitude. The metal foil, which is in intimate contact with the base plate, consolidates to the plate in a small area directly below the sonotrode. The portions of the metal foil which are not directly beneath the sonotrode do not bond to the base plate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure7-1.png", + "caption": "Figure 7 b shows the position of the edges of the blade corresponding to the convex and concave sides. We may consider that both edges of the blade contain initially the reference point P. System SP is rigidly connected to the cradle whereas system Sg is rigidly connected to the blade. System Sg is provided with two motions: i rotation about axis zg and ii translation in rotation about axis zP. The first one corresponds to the cutting motion and the second one corresponds to the motion of the cradle. System Sh is an auxiliary coordinate system wherein rotation of system Sg is considered. Location of origin Oh or Og in system SP is given by magnitudes Sr2 and q2. Axis xh is collinear with line OPOh whereas axis xg is collinear with line OgPo in an instant of the cutting motion. The cutter radius for reference point P, rc P , coincides with the cutter radius for its projection Po on coordinate plane xP ,yP . Points P1 and P2 belongs simultaneously to respective edges, the convex and the concave ones, and to the coordinate plane xP ,yP .", + "texts": [ + " Since the xis of rotation of the generating surface must be perpendicular to he closer generatrix of the root cone, different axes of rotation of he generating surfaces are needed for generation of pinion and ear. Axis zP is perpendicular to the gear root line whereas axis zF s perpendicular to the pinion root line. Both axes do not coincide ut make an angle 1+ 2, wherein i i=1,2 is the dedendum ngle. Different generating surfaces P and F are applied for eneration of gear tooth surface 2 and pinion tooth surface 1, espectively. Figures 6 b and 6 c show the gear and pinion genrator pitch cones, respectively. For derivation of gear machine-tool settings, we consider first a oint P see Fig. 7 a , which belongs to the instantaneous axis of ig. 5 Illustration of coordinate system Sg, rigidly connected o generating surface, and intersection line L of generating surace with crown gear pitch cone, in case of \u201ea\u2026 a right-hand gear nd \u201eb\u2026 a left-hand gear otation xf of pinion and gear, as a reference point of generating 01002-4 / Vol. 132, OCTOBER 2010 om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 03/17/201 surface P. Magnitude O2P is equal to the mean cone distance Am. The projection of point P on gear root line is represented by point Po and the magnitude PPo is given by b2 cos 2. The rotation of system SP represents the rotation of the cradle that carries the generating surface P. Figure 7 a shows the instant wherein p=0 and system SP coincides with system Sn. Sn is an auxiliary fixed coordinate system wherein rotation of the cradle is considered. Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use s s y i w s c w w J Downloaded Fr Initially, we can say that reference point P belongs physically to urface P. Second, we consider that point P belongs to the intantaneous axis of rotation by cutting of surfaces P and 2, that ields vP P = vP 2 3 P 2 = sin 2 cos 2 4 In case that vectors P2 and OP are parallel each other, the nstantaneous axis of rotation by cutting of surfaces P and 2 ill coincide with the instantaneous axis of rotation by meshing of urfaces 1 and 2", + " The generating surface is a cone obtained by rotation of the edge of the blade. This rotation is not related with process of generation but with the required cutting velocity. The surface parameters are u , , wherein u is the profile parameter and is the longitudinal parameter. The generating surface is defined basically with two magnitudes see Fig. 8 , the blade angle gi i=1,2 and the cutter point radius Rgi i=1,2 . Subscript i=1 is applied to the convex side and subscript i=2 is applied to the concave side. Cutter point radii are given by Rg1 =OgP1 and Rg2=OgP2 see Fig. 7 . Since the generating surface is a revolution surface, the cutter radius depends uniquely on profile parameter u as rc u = Rgi u sin gi, i = 1,2 9 wherein the upper sign is applied to the convex side and the lower sign is applied to the concave side. This criterion for the sign is considered in the following derivations. Generating surface is given in system Sh by position vector OCTOBER 2010, Vol. 132 / 101002-5 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use T M w a g b w m c l y t c 1 Downloaded Fr rh u, = u cos sin gi + cos Rgi u sin sin gi + sin Rgi \u2212 u cos gi 1 10 he unit normal is given by nh = cos cos gi sin cos gi \u2212 sin gi 11 atrix given by MPh = cos q2 \u2212 sin q2 0 Sr2 cos q2 sin q2 cos q2 0 Sr2 sin q2 0 0 1 0 0 0 0 1 12 herein q2 0 for right-hand gear and q2 0 for left-hand gear, llows to represent generating surface in system SP as rP u, = u sin gi + Rgi cos q2 + + Sr2 cos q2 u sin gi + Rgi sin q2 + + Sr2 sin q2 \u2212 u cos gi 1 13 The unit normal in coordinate system SP is given by nP = cos gi cos q2 + cos gi sin q2 + \u2212 sin gi 14 System SP rotates with the cradle the angle p", + " 23 and 24 we obtain Sr2 2 = Am cos 2 2 + rc uP 2 2Am cos 2rc uP sin 25 tan q2 = rc uP cos Am cos 2 rc uP sin 26 wherein the upper sign is applied to the right-hand gear and the lower sign is applied to the left-hand gear. Spiral angle is considered positive for the right-hand gear and negative for the lefthand gear. Relations 25 and 26 represent the values of the radial distance and the cradle angle as functions of the cutter radius of reference point P and the variables directly related with the basic data of the gear drive. Initially, point P was considered on the gear generating surface P and that is the reason why edges on both sides convex and concave intersect each other in Fig. 7 b at point P in order to make the derived relations true for both sides. But actually, the convex and concave sides are generated simultaneously on the gear tooth surface by the respective edges of the blade see Fig. 3 a . This means that edges cannot be intersected but separated through the point width of the blade. The real situation is shown in Fig. 9. However, point P is in the same position than in Fig. 7. Here is the importance of considering point P as a reference point of the head cutter. Actually, point P does not belong physically to the generating surface P but has all its motions. This means that relations derived in Eqs. 25 and 26 are still true considering rc uP =Ru, wherein Ru is the mean cutter radius see Figs. 3 a and 9 . Machine-tool settings for radial distance and cradle angle are given then by Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use w l m a s m r s a o l b t t F n J Downloaded Fr Sr2 2 = Am cos 2 2 + Ru 2 2Am cos 2Ru sin 27 tan q2 = Ru cos Am cos 2 Ru sin 28 herein the upper sign is applied to the right-hand gear and the ower sign is applied to the left-hand gear. The main cutter radius Ru has to be chosen inside the recomended interval for standard type and normally considering the vailable head cutters of the manufacturer. Figure 7 a shows system Sn and system S2. Comparison with ystems Sm2 and system Sb in Fig. 3 b allows the following achine-tool settings to be obtained. Em2 = 0 29 XB2 = 0 30 XD2 = 0 31 m2 = R2 32 The blade angles are approximated to the pressure angles at eference point P in expect of further investigation. g1 1 33 g2 2 34 The ratio of gear roll is given by see Eq. 8 m2c2 = 2 c2 = 2 p = cos 2 sin 2 35 4.2 Derivation for Uniform Type of Taper. In case of a piral bevel gear drive with uniform type of taper, the dedendum ngles 1 and 2 are equal to zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001054_s12283-010-0034-3-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001054_s12283-010-0034-3-Figure5-1.png", + "caption": "Fig. 5 These six images represent possible club designs, which would result in changes to the position of the center of mass as described in Sect. 2.2. The top-row demonstrates repositioning the center of mass along the X-axis, which would primarily affect toe-up/ down deflection (see Fig. 6). The bottom-row demonstrates repositioning the center of mass along the Y axis, which would primarily affect lead/lag deflection (see Fig. 7). The reversed condition in the top-row is analogous to hitting a ball off the face of a left-handed driver while using a right-handed swing", + "texts": [ + "2 Repositioning of the clubhead center of mass Several authors [4, 7, 8] have stated that the shaft is deflected in the lead and toe-down directions at impact because of the radial force acting on the center of mass of the clubhead, which is offset from the axis of the shaft. If the radial force is the dominating factor controlling shaft deflection at impact, then reversing the position of the clubhead\u2019s center of mass should reverse the direction of shaft bending. For example, if the center of mass is geometrically moved into a position in front of (positive Y direction) the longitudinal axis of the shaft (Fig. 5), then theoretically, the shaft should be deflected in the lag direction at impact. To help understand the influence of radial force on clubhead deflection, a series of simulations were conducted that systematically manipulated the position of the clubhead center of mass. Clubhead deflections were compared for three positions of the center of mass along both the toeup/toe-down and lead/lag axes: in-line, normal and reversed (Fig. 5). In-line refers to the center of mass being collinear with the longitudinal axis of the shaft. Normal refers to the position of the center of mass in a typical driver. Reversed refers to the placement of the center of mass in the exact opposite location along a particular axis. For example, with the \u2018normal\u2019 position, the center of mass of Club_Distal (the most distal club segment) is located 4.7 cm in the negative Y direction (Fig. 5). Therefore, the reversed condition would be to place the center of mass at 4.7 cm in the positive Y direction relative to the axis of the shaft. For each condition, the position of the center of mass of Club_Distal was only changed along either the X- or the Y-axis, and not both simultaneously. A baseline measure of clubhead deflection was established by generating an optimized simulation with the normal clubhead center of mass position first. This also permitted the identical golf swing (in terms of the \u2018golfer\u2019 portion of the model) to be used for each condition", + " Because of the offset position of the clubhead\u2019s center of mass, the shaft was gradually pulled into its maximum leading position (1.22 cm) at impact. Similarly, the shaft was gradually pulled into its maximum toe-down position (-1.33 cm) at impact when radial force acted as the lone contributor to shaft deflection (Table 1). The club showed more deflection in the toedown direction than in the lead direction because the clubhead center of mass was more offset along the X-axis (5.2 cm) than the Y-axis (4.7 cm) (Fig. 5). In the second paper of this series [2], it was shown that kick velocity peaked (7 m/s) while the clubhead was near its maximum leading position (6.25 cm). This was in contradiction to existing theory, which suggests that kick velocity should be maximized when the shaft is straight at impact [3]. However, based on the results from this study, an explanation is readily available. The influence of a large radial force (456 N), acting on the offset position of the clubhead center of mass, will continue to increase kick velocity past the neutral shaft position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000837_tmech.2008.2001689-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000837_tmech.2008.2001689-Figure2-1.png", + "caption": "Fig. 2. Position/time data as the rotor approaches a change of direction at P0 .", + "texts": [ + " Therefore, while the drive is operating with motor and load torques that are substantially constant between lowresolution position sensor inputs, the position/time relationship can be approximated by a quadratic function. A high-resolution signal is generated by approximating the input position/time data using a quadratic equation of the form shown in (3). The coefficients of the quadratic equation are derived from the low-resolution signals. If \u03b8k , \u03b8k\u22121 , and \u03b8k\u22122 are the data from sensors at times tk , tk\u22121 , and tk\u22122 , respectively, as shown in Fig. 2, then substituting into (3) yields the three equations \u03b8k = a0 + a1tk + a2t 2 k \u03b8k\u22121 = a0 + a1tk\u22121 + a2t 2 k\u22121 \u03b8k\u22122 = a0 + a1tk\u22122 + a2t 2 k\u22122 (4) which can be solved to give the quadratic coefficients a0 = D1 D a1 = D2 D a2 = D3 D (5) where D = \u2223\u2223\u2223\u2223\u2223\u2223\u2223 1 tk t2k 1 tk\u22121 t2k\u22121 1 tk\u22122 t2k\u22122 \u2223\u2223\u2223\u2223\u2223\u2223\u2223 D1 = \u2223\u2223\u2223\u2223\u2223\u2223\u2223 \u03b8k tk t2k \u03b8k\u22121 tk\u22121 t2k\u22121 \u03b8k\u22122 tk\u22122 t2k\u22122 \u2223\u2223\u2223\u2223\u2223\u2223\u2223 D2 = \u2223\u2223\u2223\u2223\u2223\u2223\u2223 1 \u03b8k t2k 1 \u03b8k\u22121 t2k\u22121 1 \u03b8k\u22122 t2k\u22122 \u2223\u2223\u2223\u2223\u2223\u2223\u2223 D3 = \u2223\u2223\u2223\u2223\u2223\u2223\u2223 1 tk \u03b8k 1 tk\u22121 \u03b8k\u22121 1 tk\u22122 \u03b8k\u22122 \u2223\u2223\u2223\u2223\u2223\u2223\u2223 . The quadratic function is evaluated continuously to give an extrapolated position estimate over the time interval beginning at tk and ending when the next position signal is generated by the low-resolution encoder. For forward motion of the rotor, as illustrated in Fig. 2, the next position signal is usually produced at the position (\u03b8k + 60)\u25e6. If the rotor first reaches a turning point (P0 in Fig. 2), however, the next low-resolution position signal appears when the rotor again arrives at the position \u03b8k . Experimental results were obtained using a novel system, as shown schematically in Fig. 3. A high-resolution incremental encoder generating 1024 pulses per revolution quadrature signals provides the reference rotor position. The encoder signals are passed to a 10-bit up/down counter, which generates a 10-bit digital word indicating the rotor position to a resolution of 0.3\u25e6. The direction of the up/down count is defined by a forward/reverse direction signal derived from the quadrature encoder signals" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure2-1.png", + "caption": "Fig. 2. The steering box \u201cscrew \u2013 nut \u2013 sector\u201d.", + "texts": [ + " 1c) relative to the articulation distance a0=OrO1 0, a0 \u00bc a0 e; a0 \u00bc m0 2 z1 \u00fe z2\u00f0 \u00de; \u00f02\u00de where m0 is the module, and z1,2 \u2013 the number of teeth. Such a gear raises two problems: the correlation of the values of geometrical parameters rb1, rb2 and e, so that the motion function \u03c62(S1) and the clearance jn between the teeth to have acceptable variation ranges; finding practical applications both in terms of functionality and variation range of the transmission ratio. An application of interest is for vehicle steering, namely the steering box of type \u201cscrew \u2013 nut \u2013 sector\u201d [7], where nut acts as a rack (Fig. 2). This type of box is suitable for a hydraulic assisted steering (for example at trucks, or tractors), but the hydraulic assistance increases the cost. In the case of some light vehicles (automobiles, small tractors), the assisted steering can be eliminated. The compensation of the actuating forces on the steering wheel, in bend, parking and other maneuvers, can be assured by increasing the gearing ratio from the middle position (rectilinear motion) towards the extreme positions of the rack, meaning variable ratio it. The classic steering box, shown in Fig. 2, carries out a constant transmission ratio. For such gearbox, an increase in the transmission ratio by 30% from the middle position towards the ends of the nut can be satisfactory, and this can be realized with the solutions shown in Fig. 1b,c. In this way, the mechanical steering gearbox can create the effect/feeling of an assisted steering system. In accordance with the intended application, the geometry of such a gear will be correlated with the construction of a steering box of type \u201cscrew \u2013 nut rack \u2013 sector\u201d (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002211_j.mechmachtheory.2012.01.020-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002211_j.mechmachtheory.2012.01.020-Figure5-1.png", + "caption": "Fig. 5. Geometrical analysis of the basic dimensions of a gear.", + "texts": [ + " Assessment of the geometrical dimensions of the gears assuming the use of a triangular file In order to confirm the above mentioned scheme, it should be proved that the root angles in all gears are equal and that there exist a fixed ratio between the number of teeth and the diameter of each gear (its \u201cmodule\u201d). It is desirable (but not necessary) that one more dimension is investigated, either the whole depth or the chord length or the arc length between the top of two successive teeth. Before proceeding with this study, we made a geometrical analysis of the basic dimensions of gears mentioned above, in order to estimate the relations between them for the case of triangular teeth. In Fig. 5 a sequence of five teeth of a gear is depicted with the associated basic dimensions. Initially the tip angle (a) and the root angle (b) are calculated as a function of the tip radius (R), the root radius (r) and the number of teeth (z): a \u00bc 2 arctan r sin \u03c0 z R\u2212r cos \u03c0 z and b \u00bc 2 arctan R sin \u03c0 z R cos \u03c0 z \u2212r : Then the root angle (b) is calculated as a function of the tip angle (a) and the number of teeth (z): b \u00bc 2 \u03c0 z \u00fe a: Finally the chord length (\u0394\u0395) is calculated as a function of the number of teeth (z) and the tip radius (R)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001843_1.39537-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001843_1.39537-Figure7-1.png", + "caption": "Fig. 7 Position trajectories of each satellite, in meters.", + "texts": [ + " The specific impulse represents the impulse per unit of propellant.When the propellantwith a higher specific impulse is used, the lesser propellant is required to provide a given amount of momentum. An unknown constant disturbance, &, is considered in the numerical simulation. The values of & are selected arbitrarily. Note that unknown disturbance is applied to only satellite A. The proposed formation flight control scheme is applied to the aforementioned mission. Figures 7\u201311 show the time histories of the position and attitude of each satellite. Figure 7 shows the trajectories of each satellite, which were the result of the first and second stages, that is, virtual structuring and virtual structure rotation. Figure 8 shows the position and attitude histories of each satellite. Each satellite moves to the desired position for construction of the virtual structure and simultaneously rotates to make the bore sight vector of the virtual structure align with the target vector. In the attitude histories of satellite A, a fluctuation due to the unknown constant disturbance can be found" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure1-1.png", + "caption": "Fig. 1 Setup of the grinding wheel: parts and assembled", + "texts": [ + " In experimental investigations it was shown that the concept of a grinding wheel with an internal cooling lubricant supply at high wheel speeds (up to 50 m/s) is possible when considering fluid dynamics [13]. This aluminum prototype wheel was designed for flow analysis only and thus was not equipped with an abrasive layer. Based on these investigations using the aluminum prototype wheel, a GIC, that is equipped with an abrasive layer, was developed. The design process of the GIC and flow analyses are presented in this paper. 3.1 Setup of the GIC The GIC is an assembly of a base plate (Fig. 1 a) and an outer cover plate (Fig. 1) that are fixed together and are equipped with an abrasive layer (electroplated single layer cubic boron nitride). Once these two components are equipped with the abrasive layer, they cannot be disassembled without destroying the abrasive layer. These two components are mounted on the machine spindle and fixed with screws. The next component is an impeller that covers the mounting-screws and assures an efficient supply of the cooling lubricant to the channels (Fig. 1c). The final component is an inner cover plate that covers the impeller and is equipped with an inlet for the coolant supply (Fig. 1d). In the following chapter the design process of the GIC will be described. 3.2 Adaption of the geometry The design of the developed GIC is based on the design of the aluminum prototype wheel presented in [13]. The aluminum prototype wheel, with a total width of 30 mm, provides channels with a height of 10 mm. The outlets of the cooling channels cover 23% of the complete circumferential surface and 70% of the circumferential surface related to the height of the channels respectively (see Fig. 2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001564_ecce.2009.5316091-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001564_ecce.2009.5316091-Figure5-1.png", + "caption": "Fig. 5: CAD-drawing of the investigated mechanical two-inertia-system", + "texts": [ + " 2 displays the structure of the setup in case of a one-inertia-system and Fig. 3 shows the measurement configuration with a mechanical twoinertia-system. In both cases the closed loop current control loop is approximated as first order lag. The experimental work was carried out with industrial servo motors equipped with deep-groove ball bearing. The data are given in the Table I. In contrast to earlier investigations the damaged bearings were in the shield of the motor and not outside. The following Fig. 4 shows the bearings of the investigated PMSM. Fig. 5 shows the arrangement for the measurements on a two inertia system. The implementation of the signal processing algorithms for calculating the frequency response data is realized on a DSP-platform. In this case a 32 bit floating point processor is utilized. The cycle time of speed and current controller can be adjusted. The minimum cycle time for the current control is 62.5 \u00b5s and for the speed controller it is 125 \u00b5s. The whole identification method is accomplished during the rest time. Thus, only the data acquisition is carried out in real time" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000299_s11044-007-9098-7-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000299_s11044-007-9098-7-Figure2-1.png", + "caption": "Fig. 2 A 3D rod sliding on a rough plane", + "texts": [ + " However, if A < 0, the singularities of rigid body model will appear, in which the mode M4 s corresponds to case of the rigid body model with no solution at all, and a so-called impact without collision occurs. By observing the ingredients of the coefficient A, it is worthy to note that the singular situations depend on the configuration of the system q and the value of coefficient of friction \u03bc as well as the slip angle \u03c6 of the contact point. In this section, let us investigate the paradoxical situation appearing in a 3D slender rod that moves on a rough surface under gravity. The rod is shown in Fig. 2 with length 2l and mass m. For simplicity, the inertial frame (i1, i2, i3) and the moving body frame (\u03be, \u03b7, \u03b6 ) initially share the same origin at point O and are combined through two successive rotations: the first rotation is counterclockwise (looking down the axis) through an angle \u03c8 about i3-axis, and the second rotation is counterclockwise (looking down the axis) through an angle \u03b8 about \u03be -axis, such that the axis of the slender rod just coincides with \u03b7-axis of the body frame. The slip angle \u03c6 is measured counterclockwise (looking down the axis of i3) from i1-axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000204_1.3002324-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000204_1.3002324-Figure7-1.png", + "caption": "Fig. 7 The effect of ball location on ultr recorded when the ball is remote from the neath the transducer, and \u201ec\u2026 the ball just pas", + "texts": [ + "5 kN, w=506rpm + W=5 kN, w=364rpm o W=10 kN, w=166rpm x W=15 kN, w=106rpm ig. 6 Reflection coefficient profile measured as the bearing otates. The center of the lubricated contact occurs at x=0. The orizontal dashed lines indicate the predictions based on Eq. 3\u2026. ournal of Tribology om: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/d 5 Discussion Under normal operation, the inner surface of the outer-raceway and the ball undergo displacement due to the contact stress. This is shown schematically in Fig. 7. The reference reflection Rref is recorded from the undeformed raceway-air interface when the ball is not in position beneath the transducer Fig. 7 a . When the ball passes underneath the transducer, the oil-film reflection R is measured, which corresponds to when the raceway is subjected to deformation by the contact Fig. 7 b . Thus the reflected signal is recorded from a deformed raceway surface, while the reference signal is from an undeformed surface. When the ball is directly below the transducer the reflection is from a symmetrical surface that has a lower radius of curvature relative than the undeformed surface. This smaller radius will cause a more divergent reflection relative to the undeformed case and hence decrease the measured reflection coefficient. When the ball is on either side of this central alignment Fig. 7 c , the reflection is from an asymmetric surface and so this will tend to skew the ultrasonic beam away from the transducer. Again, this situation would be expected to cause a reduction in the measured reflection coefficient relative to the undeformed case . This source of discrepancy is analyzed here and shown to be the cause of the anomalous reflection maximum at the contact center. The lubricated contact is elliptical and in the 6016 bearing tested, the major to minor radius aspect ratio is 11:1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001700_j.mechmachtheory.2008.12.001-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001700_j.mechmachtheory.2008.12.001-Figure1-1.png", + "caption": "Fig. 1. A prototype of 4-DOF RRPU + 2UPU PM (a) and its coordinate system model (b).", + "texts": [ + " This RRPU + 2UPU PM has some potential applications, such as parallel machine tools, sensor, surgical manipulator, tunnel borer, barbette of war ship, and satellite surveillance platform. 2. An over-constrained RRPU + 2UPU PM and its geometric constraints A 4-DOF over-constrained 3UPU PM can be created by eliminating one UPU leg from the 4-DOF over-constrained 4UPU PM [10]. Since the 4-DOF over-constrained 3UPU PM only has 3 UPU active legs, it needs the fourth actuator to drive the platform. Thus, a 4-DOF over-constrained RRPU + 2UPU PM with 3 legs (see Fig. 1) can be created from the 4-DOF over-constrained 3UPU PM by installing a rotational actuator on the base B and connecting it with one of the revolute joints of the lower universal joint. A RRPU + 2UPU PM includes a moving platform m, a fixed base B, 2 UPU legs with a linear actuator, and a RRPU (active revolute joint-revolute joint-active prismatic joint-universal joint) leg with a rotational actuator and a linear actuator. Where, m is a regular triangle with 3 vertices bi (i = 1,2,3) and 3 sides li = l and a central point o; B is a regular triangle with 3 vertices Bi and 3 sides Li = L and a central point O", + " i a lx (cm) ly (cm) lz (cm) Hi (cm) hi (cm) Si (cm2) Vi (cm3) 1 180 59 68 72 24 72 10,205 66,363 2 150 66 69 67 32 67 10,258 67,028 3 120 75 73 58 48 57 11,010 72,295 4 90 91 81 50 63 50 13,744 90,174 5 60 117 102 49 75 45 21,361 138,701 6 30 163 150 70 83 42 42,577 266,676 7 0 171 197 106 85 40 67,449 389,028 lx, ly, lz \u2013 longest distance of workspace boundary in X, Y, Z; Si \u2013 surface of 3Wi; Vi \u2013 volume of 3Wi; Hi \u2013 height of central point of lower boundary surface; h \u2013 height of the tope point of upper boundary surface. \u03b1=\u00b1150\u00ba \u03b1=\u00b1180\u00ba \u03b1=\u00b1120\u00ba \u03b1=\u00b190\u00ba \u03b1=\u00b160\u00ba \u03b1=\u00b130\u00ba \u03b1=0\u00ba X Z O r1 r2 r3 X Y Z O r1 r2 r3 B2 b1 b2 b3 B1 B3 a b c Fig. 3. A reachable workspace W of 4-DOF RRPU + 2UPU PM. (a) The isometric view. (b) The front view. (c) The tope view. Ta and the unit vector R12 are collinear. The force situation of the RRPU + 2UPU PM is similar to Fig. 1b, except that a group of vectors (v,x,vri, and xa) are replaced by (F,T,Fai, and Ta), respectively. Based on the principle of virtual work [20], the total power done by the active forces and active torque (Fa1,Fa2,Fa3,Ta) and the power done by (F,T) exerted on m at o must equal to zero. A formula for solving (Fa1,Fa2,Fa3,Ta) is derived from Eq. (9b) as follows: VT F T 1 6 \u00fe vT r Fa1 Fa2 Fa3 Ta 2 6664 3 7775 \u00bc 0) Fa1 Fa2 Fa3 Ta 2 6664 3 7775 \u00bc \u00f0J 1 e \u00de T JT F T : \u00f013\u00de 7. A reachable workspace A reachable workspace W of PM is defined as all the positions that can be reached by the central point of the moving platform" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002698_iros.2012.6385850-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002698_iros.2012.6385850-Figure2-1.png", + "caption": "Fig. 2. a) Motivation for using relative configuration in polar coordinate system. b) The robot pi re-indexes all robots based on the relative position in a polar coordinate system with pi = q0 as center.", + "texts": [ + " , i qN\u22121} contains the robots that are within a distance of 2R from pi, sorted in order of increasing distances (radii) from pi. Ties in radii are broken by ordering the robots equidistant from pi in increasing order of angles with the line joining pi and iq1, the closest robot to pi. If more than one robot lie at the shortest distance from pi, then one of these robots is randomly chosen as q1 1. For the sake of legibility, we rename robot pi as q0. An example illustrating the construction of the sets iQ is illustrated in Figure 2(b). With q0(= pi) as center, let iC1, i C2, . . . i CK , K \u2264 N\u22121, denote virtual circles of increasing radii passing through one or more robots in iQ; circle iC1 passes through iq1 and so on. Let irk be the radius of circle iCk. In the following, when the context is unambiguous, we drop the superscript i to simplify the notation and refer to iqj as qj , iCk as Ck, and irk as rk. For brevity, we use qk to refer to both the robot itself and its position. Further, by a slight abuse of notation, we use notation Ck to refer to both the virtual circle, and the set of robots on the circle Ck" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002838_j.proeng.2012.07.203-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002838_j.proeng.2012.07.203-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of a flexible displacement sensor Fig. 2. Inner construction of three port on/off valve", + "texts": [ + " In the next step, to control the McKibben artificial muscle that has a closed chamber volume, an inexpensive pressure control type quasi-servo valve using a low-cost embedded controller and a pressure transducer is proposed and tested. In addition, to improve the pressure control performance of the valve, P control method with the compensation for the supplied and exhausted flow rates is proposed and embedded into the low-cost controller. By this compensation, the linearization of the valve characteristics can be realized in addition to the improvement of the pressure control performance. Figure1 shows the schematic diagram of the proposed and tested quasi-servo valve. The quasi-servo valve consists of two on/off type control valves (Koganei Co. Ltd., G010HE-1) whose both output ports are connected each other. One valve is used as a switching valve to exhaust or supply, and another is used as a PWM control valve that can adjust output flow rate like a variable fluid resistance. The valve connected with the actuator is a two-port valve without exhaust port. The other is a three-port valve that can change the direction of fluid flow from the supply port to the output port or the fluid flow from the output port to the exhaust port" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002588_1.4005215-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002588_1.4005215-Figure3-1.png", + "caption": "Fig. 3 Cross section of the turbine simulator; P1 P5 are the plenum locations and C1 C5 are locations of the cooling air inlets or discharges. Pt and Tt denote the pressure and temperature at hot section (turbine) and Pc and Tc denote the pressure and temperature at cold section (compressor).", + "texts": [ + " Lee and Kim [37] also present a 3-D THD model for double-acting foil thrust bearings considering cooling effect of the thrust runner disc by the cooling air plenum. Most THD studies on radial foil bearings are for single pad circular AFB while many industrial foil bearings have multipad configuration. Especially, AFBs with three pads with hydrodynamic preload [9,13] as described in Fig. 1 have many attractive features such as higher rotor-bearing stability and lower start friction compared to the single pad bearings. The preload Rp in Fig. 3 is defined as the physical distance between the bearing center and the top foil pad center when the top foil is assumed to follow the ideal circular profile as attached to the bump foil. The set bore clearance CS defines the minimum clearance between the rotor and bearing at the center of each top foil. When multiple top foil pads are used, the thermal boundary condition at the leading edge of each top foil is determined from mixing behavior of cooling air with exited air from the upstream trailing edge", + " 134, MAY 2012 Transactions of the ASME Downloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jetpez/27192/ on 05/02/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use TplenumH=L \u00bc temperature of PlenumH/L (K) TR \u00bc rotor temperature (K) TsealH/L \u00bc seal temperature at high/low pressure side (K) Ttop \u00bc top foil temperature (K) T1H=L \u00bc ambient air temperatures (high and low temperature side) (K) X, Y \u00bc Cartesian coordinate for global rotor motion (or bearing in Fig. 3) z \u00bc axial coordinate ach bump \u00bc flow admittance of bump channels (kg/(Pa s)) ach LE \u00bc flow admittance of leading edge grooves (kg/(Pa s)) ach rotor \u00bc flow admittance of rotor channels (kg/(Pa s)) e \u00bc nondimensional rotor eccentricity K0 \u00bc bearing number at room temperature k \u00bc recirculation ratio l \u00bc viscosity of air (Pa s) l0 \u00bc viscosity of air at 20 C (Pa s) \u00bc Poisson\u2019s ratio of rotor material g \u00bc loss factor of bump foils h \u00bc circumferential coordinate qch bump avg \u00bc average density of air in bump channels (kg/m3) qch LE avg \u00bc average density of air in leading edge grooves (kg/m3) qch rotor avg \u00bc average density of air in rotor channels (kg/m3) qcooling \u00bc density of main cooling air at inlet (kg/m3) qdischarge \u00bc density of main cooling air at discharge to outside (kg/m3) x \u00bc rotor angular velocity (rad/s) xS \u00bc rotor excitation frequency (rad/s) [1] Agnew, G" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001937_icems.2011.6073463-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001937_icems.2011.6073463-Figure3-1.png", + "caption": "Fig. 3. Blade bending dynamics for 4-mass model", + "texts": [ + " The model system needs generator torque TG and three individual aerodynamic torques (TB1, TB2, TB3) acting on each blade. The sum of the blade torques develops the aerodynamic torque TW. It is assumed that the aerodynamic torques acting on the hub and the gearbox are zero [12, 13]. Additionally, the detailed equation of motion for 6-mass model can be found in Reference [12]. B. 4-Mass Drive Train Modeling Since the blade bending occurs at a certain distance from the joint between the blade and the hub, the blade can be split into two parts as shown in Fig. 3, the rigid blade (OA) and the flexible blade (AB) [7]. The blade sections OA1, OA2 and OA3 are collected into the hub and have the inertia HH, and the rest of the blade sections A1B1, A2B2 and A3B3 are simplified as a disk with small thickness with inertia HB about the shaft. It is assumed that the three-blade turbine has uniform weight distribution; and the turbine torque is equal to the summation of the torque acting on three blades. As a result, 4-mass model is developed from 6-mass and shown in Fig", + " Besides, the different transient behaviors between 4-mass and 3-mass II models are tiny. It means flexibilities of the high-speed shaft have little effect on transient responses. In conclusion, in terms of simulation accuracy and speed, 3-mass II model is sufficient for transient analysis in most conditions. C. Influences of flexible parameters on transient responses in 3-mass II model In order to analyze the impacts of blades flexibilities, lowspeed shaft flexibilities and inertia constant ratio of hub and blades (which stands for the A point position in Fig. 3) on the transient responses of 3-mass II model wind turbines, the following three cases are simulated in this section. A three-phase fault occurs at 0.5s and lasts for 0.25s at the output of DFIG in each case. In the first case shown in Fig. 11, the blade flexible parameter KBH is increased by 50% and 100% from the original value, respectively; while the lowspeed shaft flexible parameter KLS is increased by 50% and 100% from the original value respectively in the second case as shown in Fig. 12; finally, in the third case shown in Fig. 13, with the overall inertia of blades and hub unchanging, the hub inertia HH is increased by 50% and 100% from the primitive value, respectively, which represents the A point in Fig. 3 is closer to the blade tip. Simulation results have shown that flexible parameters and the inertia constant ratio of the blades and hub have significant influences on the transient performances. As the stiffness of blades and low-speed shaft increases, the response time is conspicuous shorter but the torque amplitude ascends in the first two cases. Moreover, in the third case, the response time becomes shorter slightly with puny amplitude decreasing, when the ratio of the rigid blade length to the flexible blade length increases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002417_iros.2011.6048792-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002417_iros.2011.6048792-Figure4-1.png", + "caption": "Fig. 4. Optimal energy density piezoelectric bimorph actuators provide mechanical power to the hip joint transmissions on HAMR3. As shown, the clamped-free cantilever produces motion into and out of the page. The clip interface at the bottom of the actuator creates a simple attachment to the circuit board body, where the three signals are provided from onboard electronics.", + "texts": [ + " This design will allow future tests on varying leg dynamics and foot attachment mechanisms. Consistent with the previous HAMR prototype and other robotic insects [11], optimal energy density piezoelectric bimorph cantilevers [12] are used for actuation on HAMR3. These actuators have proven to be suitable for the energy requirements of millimeter-scale mobile robots and have a high bandwidth, enabling quasi-static operation. The optimal energy density design, which consists of a tapered clampedfree cantilever beam, is used to minimize the actuator mass for a desired output (see Fig. 4). The nominal actuator design, composed of a central carbon fiber layer, two 125\u00b5m thick lead-zirconate-titanate (PZT5H) plates, and electrically-insulating fiberglass reinforcements, was modified from previous versions to include a mechanical clip interface and solderable connections for the three input signals (see Fig. 4). Signals are traced to the PZT plates and central electrode layer using 125\u00b5m copperclad FR-4 printed circuit board, which additionally provides a rigid mechanical ground at the base of the cantilever. Nine piezoelectric actuators drive the twelve nominal degrees of freedom on HAMR3; three actuators are associated with swing, and six with lift. The swing DOFs of each contralateral leg pair are asymmetrically coupled such that driving the actuator left moves the left leg back and the right leg forward and vice versa", + " A modular design also facilitated rapid parametric testing of linkages and actuators. The double-sided circuit board, the fabrication of which is detailed in Section III, includes bond pads for surface mount electrical components and cutouts for plug-in mechanics. The full robot assembly began by populating the circuit board with all surface mount components, and concluded with manual placement of actuators and linkages into their corresponding slot. Mechanics are fixed by sliding them backwards to engage them with the board using a simple clip mechanism (see Fig. 3 and Fig 4), assisting in alignment as well as providing a press-fit interface. Linkages were locked in place using CrystalbondT M , while actuators were fixed to their appropriate bond pads with solder. The current assembly process takes several hours to complete, an improvement from days for the previous HAMR prototypes. Although manual soldering was used in this work, the assembly method was designed to use reflow in the future for single-step assembly. The completed HAMR3 prototype is 1.7g and 48mm long and demonstrates untethered walking on flat ground" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002257_j.jsv.2011.06.018-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002257_j.jsv.2011.06.018-Figure4-1.png", + "caption": "Fig. 4. Description of contact surface in the spherical coordinates.", + "texts": [ + " Therefore, the sliding configuration is found under the action of pre-normal and friction loading and then the perturbation at the sliding equilibrium is described as shown in Fig. 3. According to the above scenario, the sliding equilibrium is to be determined first. Under the sliding state, the center of the ball is located at \u2018\u2018A\u2019\u2019 and the corresponding contact location can be described as req c \u00bc xA\u00feRer , (1) xA XAio\u00feYAjo\u00feZAko, (2) where xA is the position vector of \u2018\u2018A\u2019\u2019 at equilibrium and er is the radial direction vector of the spherical coordinates as shown in Fig. 4. The corresponding velocity vector at the contact location is given by veq c \u00bcX \u00f0Rer\u00de, (3) where the time derivative of the equilibrium deflection vector becomes zero by definition ( _xA \u00bc 0) and O is the angular velocity of the local axis (X\u00bcOko). Here (i,j,k) and (io,jo,ko) are denoted as the direction vectors of the local coordinates (x,y,z) and the reference coordinates (X,Y,Z), respectively, so that er is described as er \u00bc sinfcosyio\u00fesinfsinyjo\u00fecosfko: (4) Therefore, the sliding velocity vector of Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003412_j.issn.1004-4132.2011.04.017-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003412_j.issn.1004-4132.2011.04.017-Figure1-1.png", + "caption": "Fig. 1 The digraph composed of four agents and a leader", + "texts": [ + " Moreover, we will find in the simulation that the positive definite matrices, which make the LMIs in Theorem 1 ( or Theorem 2 ) hold for all the topologies in the switching process, are difficult to obtain for the system with switching topology. Therefore, we use a delay-dependent consensus analysis method to analyze the consensus problem of the second-order multi-agent systems with communication delay and switching topology. Example 1 The consensus under static interconnection topology. Consider a network of four agents and a leader described by (4). The interconnection topology of the agents and the leader is described in Fig. 1. Obviously, the static leader is the globally reachable node of the topology. We adopt the time-varying communication-delay \u03c4(t) = 0.3t satisfying Assumption 1, i.e., \u03c1 = 0.3. We choose \u03b31 = 0.01, \u03b32 = \u03b33 = 0 and \u03b34 = 0.02, and assume k1 = k2 = . . . = kn = k for simplicity. By using the LMI toolbox in Matlab for the LMI (7) in Theorem 1, the minimum control gain k, which guarantees that the LMI (7) holds, is kmin 0.217. Choosing k > kmin = 0.217, we obtain that each agent converges to the leader\u2019s states with the initial states of the agents generated randomly and the static leader\u2019s position chosen as x0 = 3 (see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001209_j100593a043-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001209_j100593a043-Figure2-1.png", + "caption": "Figure 2. Reaction cell used with Picker and GE x-ray diffractometers.", + "texts": [ + " The effect of buoyancy was determined by calibrating in air and in vacuo and the deviation was found to be 0.05 mm in the total extension of 3 X IO2 mm at 1 atm of air relative to vacuum. The fraction of a phase present in a powdered mixture may be determined from the integrated intensity of the corresponding x-ray diffraction lines,8cid For powders with pd << 1 (p IO3 cm-l is the linear absorption coefficient and d is the normal dimension) the rate of change of the latter is then a measure of the rate of change of the corresponding phase. The reaction cell shown in Figure 2 was constructed for use with a GE XRD-6 x-ray diffractometer. The sequence of measurements involved (a) measurement of the x-ray spectrum of a blank cell (using a CA-8-S/Cu tube with a 0.4O slit and a Ni filter) from 20 = 3 to 80\u00b0 a t the highest spectrometer sensitivity, (b) loading and pumping of cell to 10-5 Torr overnight, (c) measurement of reactant spectrum in same range as in (a), (d) starting the reaction by letting gas flow into the cell from an ampule or from a gas reservoir and measuring the time intervals with the synchronous motor of the recorder, and (e) scanning from 20 = 15 to 8 O and back a t the rate of 2O/min after earlier runs showed there were no base line changes from 3 to go for both gases investigated within the first 3 min of the reaction", + " In powders (2H-TaS2, d, << d, < 38 pm) the rate of disappearance of the integrated intensity of the 100 x-ray reflection ( y h ~ ) is nearly equal to that for the 004 reflection of the reactant phase. Insertion of NH3 and N2H4 into Layer Disulfides 8 00 , , , , 3007 -- 0 2 0 -- 0 10 0 00 X O x m - I n ( l - - ) I 1 I . 4 063 D-23 3 0 . i X O 1 0 0 9 0 8 07 0 6 0 5 0 4 0 3 I I /,Ioc 020J Flgure 10. Equilibrium data (isotherms) for 2H1-TaS2-(NH3),, complexes. The presence of hysteresis Is indicated In the insert. Two single crystals (2H-TaS2) of approximately equal weight (19 mg) and da2 - 0.2-0.1 cm2 were allowed to react in separate cells as shown in Figure 2 with NzH4 (PI = 5 Torr a t room temperature). One crystal was aligned with the c axis normal to the Be window, the other with the a axis normal to the window. For the former only the 001 lines can be measured and the initial rate of disappearance of the integrated intensity for the 002 line is -(l/yo02) (dy002l dt = 0.02 min-'. For the latter only the hOO lines could be measured. The 100 and 200 lines had sufficient intensity but the 100 line was selected. The spectrometer was adjusted for 28 (maximum intensity of 100) = 31" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure14-1.png", + "caption": "Fig. 14 Obtained geometry for \u201ea\u2026 the gear and \u201eb\u2026 the pinion of the spiral bevel gear drive 20\u00c343", + "texts": [ + " 13 For determination of point width: \u201ea\u2026 illustration of uxiliary coordinate system St and \u201eb\u2026 illustration of arc M1M2 \u0302 ontaining point N urface. The basic data of the gear drive are shown in Table 2. The ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 03/17/201 gear machine-tool settings obtained by application of the developed theory considering duplex taper are represented in Table 3. The point width was obtained considering =0. The obtained geometry for the gear of the spiral bevel gear drive 20 43 is shown in Fig. 14. This geometry is obtained from the determined gear machine-tool settings represented in Table 3. Table 4 shows the pinion machine-tool settings obtained from computations following the approach described in Ref. 8 , based on local synthesis, which provides a longitudinally oriented path of contact and a predesign function of transmission errors of 8 arcsec of maximum level. Figure 15 shows the results of TCA for the optimized design of the spiral bevel gear drive 20 43. Figure 15 a shows the bearing contact for the concave side of the gear tooth surface and Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000432_s11071-008-9400-0-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000432_s11071-008-9400-0-Figure1-1.png", + "caption": "Fig. 1 The rub and impact forces", + "texts": [ + " Equation (2) can then be reduced to \u22022p \u2202z2 = \u22126\u03bc\u03c9c\u03b5 sin \u03b8 + 12\u03bc(c\u03b5\u0307 cos \u03b8 + c\u03b5\u03d5\u0307 sin \u03b8) \u03be(h, l)Gz , (3) with boundary conditions \u23a7 \u23aa\u23a8 \u23aa\u23a9 \u2202p \u2202z = 0, z = 0, p = 0, z = \u00b1L 2 , such that p = \u2212 3\u03bcc \u03be(h, l)Gz [ (\u03c9 \u2212 2\u03d5\u0307)\u03b5 sin \u03b8 \u2212 2\u03b5\u0307 cos \u03b8 ] The resulting bearing forces about the journal center in the radial and tangential directions are determined by integrating (4) over the area of the journal sleeve: fr = \u222b \u03c0 0 \u222b L 2 \u2212 L 2 pR cos \u03b8 dz d\u03b8, (5) ft = \u222b \u03c0 0 \u222b L 2 \u2212 L 2 pR sin \u03b8 dz d\u03b8, (6) fe = \u2212fr, f\u03d5 = \u2212ft . Therefore, fe = \u2212\u03bcL3R 2c2 \u222b \u03c0 0 { [(\u03c9 \u2212 2\u03d5\u0307)\u03b5 sin \u03b8 \u2212 2\u03b5\u0307 cos \u03b8 ] cos \u03b8 Gz[(1 + \u03b5 cos \u03b8)3 \u2212 12(l\u2217)2(1 + \u03b5 cos \u03b8) + 24(l\u2217)3 tanh( 1+\u03b5 cos \u03b8 2l\u2217 )] } d\u03b8, (7) f\u03d5 = \u2212\u03bcL3R 2c2 \u222b \u03c0 0 { [(\u03c9 \u2212 2\u03d5\u0307)\u03b5 sin \u03b8 \u2212 2\u03b5\u0307 cos \u03b8 ] sin \u03b8 Gz[(1 + \u03b5 cos \u03b8)3 \u2212 12(l\u2217)2(1 + \u03b5 cos \u03b8) + 24(l\u2217)3 tanh( 1+\u03b5 cos \u03b8 2l\u2217 )] } d\u03b8. (8) 2.2 Rub-impact force Figure 1 shows the radial impact force f1 and the tangential rub force f2.f1 and f2 could be expressed as [22] f1 = (e \u2212 \u03b4)kc, (9) f2 = (f + bv)f1, if e \u2265 \u03b4. (10) Then we could get the rub-impact forces in the horizontal and vertical directions: Rx = \u2212 (e \u2212 \u03b4)kc e [ X \u2212 (f + bv)Y ] , (11) Ry = \u2212 (e \u2212 \u03b4)kc e [ (f + bv)X + Y ] . (12) 2.3 Dynamic equations Figure 2 shows a flexible rub-impact rotor supported by two couple stress fluid film journal bearings in parallel with nonlinear springs. Om is the center of rotor gravity, O1 is the geometric center of the bearing, O2 is the geometric center of the rotor, O3 is the geometric center of the journal, m is the mass of the rotor, m0 is the mass of the bearing housing, Ks is the stiffness of the shaft, K1 and K2 the stiffness of the springs which support the bearing housings, C1 and C2 the damping coefficient of the supported structure and the quadratic damping of the rotor disk, respectively, \u03c1 is the mass eccentricity of the rotor, \u03c6 the rotational angle, R the inner radius of the bearing housing and r the radius of the shaft" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003080_coase.2012.6386340-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003080_coase.2012.6386340-Figure11-1.png", + "caption": "Fig. 11. Experimental results of shape recognition: (a) round peg, and (b) square peg.", + "texts": [ + " The diameter of the round peg is 50 mm, and each edge of the square peg is 50 mm. A relatively large force of 20 N was imposed to the peg to maintain a stable contact between the peg and the hole. Since the maximum radius of the peg is 25 mm, the imposed moment was 1.0 Nm. 10,000 data points were collected to extract the contour of the peg, which took approximately 10 s to complete. Then, the contour of the peg was found by computing the dominant points and plotting them in polar coordinates. The results of the shape recognition experiments are illustrated in Fig. 11. Fig. 11(a) shows the result of the shape recognition of the round peg, and the maximum recognition error was found to be 5%. The result for the square peg is shown in Fig. 11(b), with the maximum recognition error of 8%. It follows from the results that the shape of each peg was successfully estimated using the proposed algorithm. Several experiments were conducted to investigate the performance of the hole detection algorithm under the conditions of only lateral error, only angular error, and combined errors, as shown in Fig. 12.The tolerance between the hole and the peg was 0.1 mm. Figure 13 shows the experimental results of hole detection with respect to the conditions shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000531_9783527627059.ch1-Figure1.20-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000531_9783527627059.ch1-Figure1.20-1.png", + "caption": "Figure 1.20 Assembly of Au-nanoparticle-reconstituted GOx electrode by adsorption of Au-nanoparticle-reconstituted GOx onto a dithiol monolayer and a stepwise build-up of the electrode by deposition of the Au-nanoparticle FAD onto the dithiol surface followed by the reconstruction of the apo-GOx.", + "texts": [ + "1 Plugging Nanomaterials into Proteins \u2013 Nanoparticles The modification of electrodes with nanoparticles followed by the attachment of redox enzymes is one approach to nanostructuring electrodes that has been successful at achieving direct electron transfer to enzymes. Perhaps the most stunning example of this approach is to use nanoparticles to wire into glucose oxidase, as has been achieved by Willner and coworkers [87]. In this work, a gold electrode was modified with a dithiol self-assembled monolayer such that one thiol attached to the gold electrode and the other to gold nanoparticles, which were 1.4 nm in diameter. Active glucose-oxidase-modified gold nanoparticles were produced in one of two ways (Figure 1.20). In the first the redox-active center of glucose oxidase, flavin adenine dinucleotide (FAD), was immobilized onto a SAM-modified nanoparticle that was subsequently attached to the electrode. The active enzyme was produced by reconstitution of the apo-enzyme around the FAD-modified nanoparticle attached to the electrode surface. In the second approach, the enzyme was reconstituted onto the FAD-modified nanoparticles in solution prior to attaching the nanoparticle to the electrode surface. Both strategies produced active enzyme and enzyme electrodes where direct electron transfer to glucose oxidase could be achieved with almost identical performance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure2-1.png", + "caption": "Fig. 2 SHF-type SWG", + "texts": [], + "surrounding_texts": [ + "Table 1 shows the requirements of SWG for this development. Considering that the major application of SWG for space applications is a paddle-drive and an antennapointing mechanisms, the SHF-type SWG with a hollow-shaped shaft was selected as a product for the development (Figs 2 and 3). It is made of stainless steel to prevent rust (Table 2) and its size is 20 (catalogue model number), which is currently the most used for space applications. Its reduction ratio is 160:1, which is the largest ratio among standard ratios and was selected because it is generally operated at very low The teeth of circular spline and flexspline, between the inner part of flexspline and outer part of wave generator and between races and balls of wave generator bearing, is lubricated by a multiply alkylated cyclopentane (MAC) grease. The bearing ball separator of wave generator is made of cotton-based phenolic resin, impregnated with MAC oil in vacuum. The grease application part and the quantity are presented in Table 3." + ] + }, + { + "image_filename": "designv11_3_0001250_iecon.2009.5415267-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001250_iecon.2009.5415267-Figure1-1.png", + "caption": "Fig. 1. Vehicle schematic for vertical flight mode of Bidule", + "texts": [ + " Tail-sitters are therefore very useful when there are no facilities for taking-off and landing for standard airplanes. One of the challenges is to control the aircraft during the taking-off and landing phase. This paper presents the tail-sitter prototype developed jointly by the University of Sydney and the University of Compiegne. In order to increase the magnitude of the couples to control the attitude the vehicle has two Variable Pitch Propellers (VPP). The vertical fixed wing tailless aircraft with two propellers is shown in Figure 1. Bidule-CSyRex is meant to have two modes of operations, hover and forward flight. In hover mode, a vertical airframe attitude is adopted during take-off, hover and landing while in forward flight, a horizontal airframe attitude during cruise is maintained, just like conventional airplanes. Tail-sitters have a complex flight dynamics in hover mode, making them typically very difficult to control. Robust control has been intensively studied resulting in several approaches, including H\u221e approach [11], [27]; robust control with parametric uncertainty in [1], [4], [15], and also robust stability analysis with parametric uncertainty for linear time-delay systems [20]", + " The transformation of the components of the angular velocity generated by a sequence of Euler rotations from the body to the local reference system can be written as: H (\u03a6) = \u23a1 \u23a2\u23a3 1 t\u03b8s\u03c6 t\u03b8c\u03c6 0 c\u03c6 \u2212s\u03c6 0 s\u03c6/c\u03b8 c\u03c6/c\u03b8 \u23a4 \u23a5\u23a6 (4) where s, c and t are used to denote the sin, cos and the tan respectively, then using (4), the kinematic equations (1) can be rewritten as: \u03c6\u0307 = P + tan \u03b8 (Q sin \u03b8 +R cos\u03c6) (5) \u03b8\u0307 = Q cos\u03c6\u2212R sin\u03c6 (6) \u03c8\u0307 = (Q sin\u03c6+R cos\u03c6) / cos \u03b8 (7) The term Jb in (2) represents the inertia matrix, and is defined by Jb = \u23a1 \u23a2\u23a3 Jx Jxy Jxz Jyx Jy Jyz Jzx Jzy Jz \u23a4 \u23a5\u23a6 If the Bidule CSyRex is assumed to have the body axis xzplane coincident with the plane of symmetry, then the product of inertia Jxy and Jyz vanish. This tail-sitter configuration, also presents a plane of symmetry in the xy-plane, then the product of inertia Jxz = 0. Then Jb and its inverse can be written as Jb = diag(Jx, Jy, Jz) and (Jb)\u22121 = diag( 1 Jx , 1 Jy , 1 Jz ). Note that the mass of the elevons is neglected. The aerodynamics and thrust moments can be denoted by Mb A,T =[ m n ]T , they are shown in the Figure 1, then using the matrix of inertia and the moment vector, (2) yields: P\u0307 = (Jy \u2212 Jz)QR Jx + Jx (8) Q\u0307 = (Jz \u2212 Jx)RP Jy + m Jy (9) R\u0307 = (Jx \u2212 Jy)PQ Jz + n Jz (10) Since lateral, longitudinal and axial dynamics will be analyzed separately, the following assumptions are needed: Assumption 1: For lateral dynamics it is assumed that the pitch and yaw rates are zero. Assumption 2: For longitudinal dynamics, the vehicle is considered to be a tailless aircraft flying in forward flight. Assuming that the roll angle is small enough and the roll rate is zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002184_cp.2012.0275-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002184_cp.2012.0275-Figure9-1.png", + "caption": "Figure 9: Magnet shape with optimal third order harmonic.", + "texts": [ + "0/1k (15) As stated above, when the third harmonic is introduced into the magnet shape, the airgap length for inverse cosine and magnet height for sinusoidal magnet shape can be respectively expressed as: )3sin()sin( a l l gd g (16) and )3cos()(cos( papmh (17) where a is the amplitude of third harmonic injecting into the magnet shape. The finite element results for the slotless machines are given in Figures 7 and 8 for inverse cosine shape airgap and sinusoidal shape magnet methods, respectively. As can be seen, the optimal amplitude of 3rd harmonic for inverse cosine and sinusoidal magnet shape is always 1/6 of that of the fundamental. In order to produce the required 3rd order flux density harmonic, the magnet shape needs to be modified as shown in Figure 9 (a) and (b) for inverse cosine shape airgap and sinusoidal shape magnet methods, respectively, while their airgap field distributions and harmonic contents are shown in Figures 10 and 11, respectively. When the 3rd order harmonic is introduced, (third harmonic 1 in Figure 9), both the magnet thickness can be reduced whilst the magnet material assumption is compared in Table 3, but the average torque is kept almost the same, Figure 13 (a). The airgap is actually increased by introducing the 3rd order harmonic, which is beneficial for manufacturing. There are two ways of keeping the same airgap length: one is to increase the rotor back-iron radius (third harmonic 1+) and another is to increase the magnet overall thickness (Third harmonic 2). Third harmonic 1 and third harmonic 1+ have the same magnet volume, but the latter method can result in higher torque due to reduced effective airgap length" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001629_285710-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001629_285710-Figure1-1.png", + "caption": "Fig. 1. Schemat'ic diagram of 4~ plastic p-scintillation counter.", + "texts": [ + " Smith, Seliger and Steyn (1956) and also Mann and Seliger (1956) have described a 4n p-scintillation counter using two anthracene crystals viewed by two photomultipliers, working either in addition or in coincidence. For experiments reported here a simplified 4n scint'illation system has been developed in which two thin plastic scintillators, enclosing the source, are viewed by a single cooled photomultiplier. 0 2 . COUPJTER DESIGN AND CHARACTERISTICS The plastic scintillators used were type NE 102 (Nuclear Enterprises Ltd.). One face of the pair forming the scintillator sandwich was mounted on the end window of a photomultiplier (13 st'age 95148 E.JI.1. Ltd.) as shown in fig. 1. Pulses from the photomultiplier were fed through a pre-amplifier to a linear amplifier discriminator unit (type 652, Isotope Developments Ltd.) and a scaling unit or rat'emeter with input resolving time of 5 psec. Preliminary experiments were carried out with pairs of scintillators each 1 in. or 1.5 in. square (2.5 or 3.8 cm) and of thicknesses 0.026, 0-035, 0.05, 0.126 or 0.250 in. (0.64, 0.89, 1.3, 3 . 2 : 6.4 mm). Using a z04T1 source (fig. 2 a ) i t was found that the slope of a plot of counting rate versus discriminator bias potent'ial, and also the background counting rat'e, increased as the thickness of the scintillator was raised from 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000233_tie.2007.898297-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000233_tie.2007.898297-Figure2-1.png", + "caption": "Fig. 2. MEMS-based pneumatic microactuator array for air-flow distributed micromanipulation.", + "texts": [ + " The basic architecture of an FPGA consists of a large number of configurable logic elements and a programmable mesh of interconnections. New architectures of FPGAs both lower the cost and improve functionality, enabling them to cost-effectively compete with ASICs (even for high fabrication volume density). This technology has been used for motion control [17], [18], and is ready to be applied to new applications as MEMS, particularly for embedded and distributed devices [19]. In this paper, the authors studied an FPGA-based decentralized control system for a MEMS-based distributed airflow micromanipulator device. Fig. 2 illustrates the MEMS chip as reported [20]. First, we focused on the analysis of a class of decentralized decision-making schemes for the control method of the distributed system [21]. The decentralized decision-making is appropriate control architecture for distributed agents. Several different hardware architectures were tested to make the FPGA resource amount as low as possible, and the controller/manager as simple and efficient as possible. In Section II, a brief approach of the MEMS chip fabrication process is given" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003237_s11665-013-0583-2-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003237_s11665-013-0583-2-Figure15-1.png", + "caption": "Fig. 15 Typical results pertaining to the spatial distribution of the temperature field in the weld region over the (analyzed) right-portion of the MIL A46100 weldment at welding times of: (a) 0.6 s; (b) 2.1 s, (c) 3.6 s; and (d) 4.8 s under the following welding conditions: welding open-circuit voltage = 30 V, welding current = 200 A, electrode diameter = 1 mm, electrode-tip/weld distance = 1.3 cm, electrode feed-rate = 10 cm/s, and gun travel speed = 1 cm/s", + "texts": [ + " For example, the results pertaining to the spatial distribution of the equivalent plastic strain and the residual von Mises stress in the weldment could be potentially quite important relative to the overall functional and mechanical performance of the GMAW joint. On the other hand, detailed results pertaining to the spatial distribution and temporal evolution of the temperature within the FZ and the HAZ (and their dependence on the GMAW process parameters) are the key input to the computational analysis dealing with the prediction of the material microstructure and property distributions within the weld region. 3.1.1 Temporal Evolution of the Weldment Temperature Field. Figure 15(a) to (d) shows typical results pertaining to the temporal evolution of the temperature field within the weld region over the (analyzed) right-portion of the weldment. The results displayed in Fig. 15(a) to (d) are obtained at relative welding times of 0.6, 2.1, 3.6, and 4.8 s, respectively, and for the following selection of the GMAW process parameters: welding open-circuit voltage = 30 V, welding current = 200 A, electrode diameter = 1 mm, electrode-tip/weld distance = 1.3 cm, electrode feed-rate = 10 cm/s, and gun travel speed = 1 cm/s. To improve clarity, regions of the weldment with a temperature exceeding the liquidus temperature are denoted using red. Examination of the results displayed in Fig. 15(a) to (d) reveals that: (a) the FZ, after a brief transient period, acquires a nearly constant size and shape, as it moves along the welding direction (to track the position of the weld gun); (b) as welding proceeds, natural convection, and radiation to the surroundings, together with conduction through the adjacent work-piece material region, cause the previously molten material within the FZ to solidify (and to continue to cool); and (c) under the given welding conditions, the FZ extends downward by approximately 40-45% of the workpiece thickness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003725_icuas.2015.7152366-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003725_icuas.2015.7152366-Figure6-1.png", + "caption": "Fig. 6. SabRe (Soufflerie bas Reynolds) closed-loop wind tunnel facility.", + "texts": [ + " For the sake of completeness, this section superficially describes the wind tunnel campaign that supported this work (see [11] for more information) and how the longitudinal subset of the collected data was used to interpolate the aerodynamic coefficients, tune the interaction factor ki and validate the model. Propulsion identification was carried out in [12] and the respective parameters (among all others identified in this work) can be found in the appendix section. The experiments were ran at the SabRe closed-loop wind tunnel (Fig. 6) located at ISAE and capable of delivering low Reynolds stable and uniform flow at a wind velocity range of 2 to 25 m/s, thus ideal for experimenting full-span micro air vehicles. Although a 6-component study was performed (3-dimensional forces and moments), this paper focus only on longitudinal quantities, i.e., drag D, lift L and pitching moment M measured with zero sideslip. Forces and moments were measured by means of a calibrated 5-component internal balance in two different configurations (see figure 7) in order to obtain the 6 force/moment components (see [11] for more information)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003189_j.cnsns.2011.04.012-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003189_j.cnsns.2011.04.012-Figure4-1.png", + "caption": "Fig. 4. vertical", + "texts": [], + "surrounding_texts": [ + "The geometric details of a four-lobe noncircular bearing configuration are shown in Fig. 1. Analysis of aerodynamic four-lobe bearing involves solution of the governing equations separately for an individual lobe of the bearing, treating each lobe as an independent partial bearing. To generalize the analysis for all noncircular geometries, the film geometry of each lobe is described with reference to bearing fixed Cartesian axes (Fig. 1). Thus, the film thickness in the clearance space of the kth lobe, with the rotor in a dynamical state is expressed as\nh \u00bc C \u00f0Xj\u00de cos h \u00f0Yj\u00de sin h\u00fe \u00f0C Cm\u00de cos h hk 0\n\u00f01\u00de\nwhere \u00f0Xj;Yj\u00de is the rotor center coordinate in the dynamical state and hk 0 is angle of lobe line of centers. C and Cm are conventional radial and minor clearances, when journal and bearing geometric centers are coincident. The pressure governing equation of isothermal flow field in a bearing lobe is modeled by the Reynolds equation as follow [11]\n@\n@X h3P\n@P\n@X\n( ) \u00fe @\n@Y h3P\n@P\n@Y\n( ) \u00bc 6 l U @\n@X \u00fe 2\n@\n@ t\n\u00f0P h\u00de \u00f02\u00de\nin which P is the absolute gas pressure, l is the gas viscosity, U is the peripheral speed of the rotor and t is the time. It will be more convenient to express P as:\nP \u00bc Pa \u00fe P \u00f03\u00de\nwhere Pa and P are the ambient and partial pressure, respectively. In order to non-dimensionalize equations (1) and (2), let\nCm \u00bc Cd; \u00f0Xj;Yj\u00de \u00bc Cm\u00f0Xj;Yj\u00de; X \u00bc Rh; Y \u00bc Rn; h \u00bc Cmh; P \u00bc PaPU \u00bc U0U \u00bc R x0U; t \u00bc s x0\nTrajectory of the rotor center at mr =20.7 kg (a); phase portraits of rotor center (b) and power spectra of rotor displacement in horizontal (c) and (d) directions for k = 1.", + "that x0 is the rotational speed in the steady state and R is the rotor radius. Substituting these variables in Eq. (1), the nondimensional film thickness can be obtained as:\nh \u00bc 1 d \u00f0Xj\u00de cos h \u00f0Yj\u00de sin h\u00fe 1 d 1 cos h hk 0\n\u00f04\u00de\nand by substituting Eq. (3) in Eq. (2) and simplifying, the Reynolds equation in non-dimensional form can be expressed as\n@ @h h3\u00f0P \u00fe 1\u00de @P @h \u00fe @ @n h3\u00f0P \u00fe 1\u00de @P @n \u00bc K U @ @h \u00fe 2 @ @s f\u00f0P \u00fe 1\u00dehg \u00f05\u00de\nwhere h and n are the coordinates in the circumferential and axial directions, respectively and K \u00bc 6 l x0R2\nPaC2 m is the dimensionless parameter called the compressibility number or bearing number. The Reynolds equation is a nonlinear partial differential equation, therefore, can be solved using finite element method. For this purpose, let the function variable\nW \u00bc W\u00f0s\u00de \u00bc Ph\nbe introduced into Eq. (5) which then becomes\n@ @h h\u00f0W\u00fe h\u00de @W @h \u00f0W\u00fe h\u00deW @h @h \u00fe @ @n h\u00f0W\u00fe h\u00de @W @n \u00bc K U @ @h \u00fe 2 @ @s \u00f0W\u00fe h\u00de \u00f06\u00de\nFor the finite element formulation, the Galerkin\u2019s weighted residual of Eq. (6) for an element of the discretized space domain of W field is written as\nR R\nAe @We @t 1 2K @ @h h\u00f0We \u00fe h\u00de @We @h \u00f0W e \u00fe h\u00deWe @h @h\n1 2K @ @n h\u00f0We \u00fe h\u00de @We @n\nn oh \u00fe 1\n2 U @ @h \u00f0W e \u00fe h\u00de \u00fe @h @s Ne i dhdn \u00bc 0 \u00f07\u00de\nTrajectory of the rotor center at mr =23.2 kg (a); phase portraits of rotor center (b) and power spectra of rotor displacement in horizontal (c) and (d) directions for k = 1.", + "where Ne i is an approximation function and Ae is the area of the element. By considering the discretized domain of W variable and let, in an element \u2018e\u2019, the W function be approximated as\nWe \u00bc Xne\nj\u00bc1\nNe j Wj\u00f0s\u00de \u00f08\u00de\nin which \u2018e\u2019 refers to an element, ne is the number of nodes in the element, Ne j \u2019s are the shape functions and Wj\u2019s are the nodal values of W at time s. Applying Eq. (8) in to Eq. (7) and with some integral simplification, the finite element equations for an element of the discretized flow filed domain can be obtained as\n\u00bdF ef _Wge \u00bc fVge \u00fe fQge \u00f09\u00de\nin which the components of the element matrices are\nFe ij \u00bc Z Z Ae Ne i Ne j dhdn \u00f010:a\u00de\nVe i \u00bc 1 2K Z Z Ae \u00f0We \u00fe h\u00de h @We @h @Ne i @h \u00fe @W e @n @Ne i @n\nWe @h\n@h \u00feKU\n@Ne\ni @h dhdn Z Z Ae @h @s Ne i dhdn \u00f010:b\u00de\nQe i \u00bc Z se \u00f0We \u00fe h\u00de h @We @h We @h @h KU Ne i dn\u00fe Z se \u00f0We \u00fe h\u00deh @W e @n Ne i dh \u00f010:c\u00de\nwhere Seis the boundary of the element.\nTrajectory of the rotor center at mr =29.9 kg (a); phase portraits of rotor center (b) and power spectra of rotor displacement in horizontal (c) and (d) directions for k = 1." + ] + }, + { + "image_filename": "designv11_3_0001117_045104-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001117_045104-Figure5-1.png", + "caption": "Figure 5. Measurement principle of the tooth profile form of an involute gear.", + "texts": [ + " In the case of calibrating the gear checker using such a noninvolute helicoid artefact, the theoretical measurement results of the artefact, in other words, the calculated measurement results assuming the use of a gear checker without any error factor and the actual measurement result using a real gear checker with some error factors are compared. The difference between them is taken as the calibration result. The theoretical measurement result varies according to the form checking ball radius (cf figure 4) and the distance between the balls of the ball artefact. The VGC enables the virtual measurement of an arbitrary-shaped object such as a ball artefact, and the theoretical measurement result can be calculated easily. The bold line in figure 5 shows the base circle (radius rb) and the corresponding involute curve. The involute curve and line AB (line of action) intersect at point T. Assume that the base circle and the involute curve rotate as one body by \u03b8 and then go to the position indicated by the dotted line in figure 5. In this state, the intersection between the involute curve and line AB is at point S. The distance between points T and S is as follows: TS = rb\u03b8. (2) If this curve has a form deviation from the involute curve, the distance TS is not identical to rb\u03b8 . The difference between the distance TS and rb\u03b8 denotes the profile deviation. A tooth profile deviation curve is typically expressed as a graph with the profile deviation on the y-axis and the rotational angle of the base circle on the x-axis. There are various methods of measuring the profile deviation [1]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.33-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.33-1.png", + "caption": "Fig. 2.33 The CAD model of the Swiss lever escapement", + "texts": [ + "32. Appearing in the middle of the nineteenth century in Switzerland, the Swiss lever escapement is a modification of the English lever escapement. It is not clear who invented the Swiss lever escapement, though it was probably a team effort. It has been the most commonly used escapement in the world ever since. In fact, at least 98% of the existing mechanical movements use this escapement because of its high degree of accuracy and reliability. A model of the Swiss lever escapement is shown in Fig. 2.33. Similar to the English escapement, it consists of a balance wheel, hairspring, pallet fork and escape wheel. The pallet is shaped like a fork, giving it the name \u2018\u2018pallet fork.\u2019\u2019 The pallet fork results in two significant improvements over the English lever escapement. First, the centers of the balance wheel, the pallet fork and the escape wheel are aligned in one line, making the power transmission more efficient and stable. Second, the pallet fork needs only to swing a small degree (around 50 and 25 on each side) and hence, does not get much nonlinearity. Consequently, the accuracy of the timekeeping is being improved. The Swiss lever escapement has had a number of different versions. For the model in Fig. 2.33, the escape wheel has 15 club teeth; therefore, the angle for each impulsive movement is 360/(2 9 15) = 12. Here, the factor 2 is resulted from the swinging of the balance wheel. The operation of the Swiss lever escapement is somewhat similar to the English lever escapement. Because of its significance, we developed its mathematical model step by step, as detailed in Chap. 3. For the purpose of demonstration, a computer animation is shown on the Springer Website http://extra.springer.com/ 2012/978-3-642-29307-8" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001741_j.commatsci.2009.02.007-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001741_j.commatsci.2009.02.007-Figure7-1.png", + "caption": "Fig. 7. Finite element models for a freestanding strut model taken off from WBK truss; (a) the initial shape, (b) the deformed shape under compression with ball jointed ends, (c) the shape with fixed ends.", + "texts": [ + " It was assumed that the relation, P \u00bc ffiffiffi 6 p F, was valid between the externally applied compressive load, P, and the force applied to a single strut in the unit cell, F, and that another relation, dP \u00bc 10 9 ffiffiffi 6 p dF , was valid between displacement dP occurring at the loading point due to P, and displacement dF occurring in the single strut due to F (see Appendix). And the force\u2013displacement relation of the single strut, F dF, was estimated from P dP. In addition, to verify the boundary conditions at the brazed joint ends, we compared F dF estimated from P dP with the finite element results of a freestanding strut under compression (e = 13%) for two extreme boundary conditions (ball jointed and fixed ends). Fig. 7a shows the initial configuration of the strut, and Fig. 7b and c show its deformed shapes under two extreme boundary conditions. For the given conditions, the force\u2013displacement of single strut was 1 For interpretation of color in Figs. 9,11,14 and 15, the reader is referred to the web version of this article. determined and utilized to get the moduli for the next analysis on the bulk WBK of many cells. Fig. 8 shows the force\u2013displacement curves for the three different conditions. Here, the curve with open circles was estimated from the load\u2013displacement relation for the unit cell model of WBK, and the remaining curves were obtained for the freestanding strut models with two different boundary conditions (ball jointed and fixed ends)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001838_tmag.2009.2012780-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001838_tmag.2009.2012780-Figure1-1.png", + "caption": "Fig. 1. Flux tubes for a half part of the studied machine.", + "texts": [ + " Current version published February 19, 2009. Corresponding author: A. Mahyob (e-mail: amin.mahyob@unilehavre.fr). Color versions of one or more of the figures in this paper are available online at http://ieeexplore.ieee.org. Digital Object Identifier 10.1109/TMAG.2009.2012780 time compared to the FEM. This paper focuses on the modeling and simulation of stator inter-turn fault in induction machines based on permeance network method. An electrical machine can be represented as a set of flux tubes (Fig. 1) characterized by their magnetic permeances. These permeances are expressed as functions of the machine geometry and the instantaneous fluxes flowing in each one of them [5]\u2013[8]. Exploiting the flux tubes of Fig. 1, one can deduce the magnetic equivalent circuit of the induction machine as shown in Fig. 2 where one can distinguish the stator magnetic circuit region, the air-gap region as well as the rotor magnetic circuit region. The stator slot currents are modeled by magnetomotive force (m.m.f.) sources in series with the tooth permeances [6], [7]. The relation between these m.m.f. and the phase currents can be given by the following compact matrix form: (1) where is the vector of tooth m.m.f., is the vector of the phase currents and is the matrix that relates the tooth m" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002625_icara.2011.6144876-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002625_icara.2011.6144876-Figure1-1.png", + "caption": "Fig. 1. Six degrees of freedom of an AUV", + "texts": [ + " The first issue is system identification of AUV dynamics to obtain the coupled six degree of freedom, nonlinear dynamic model. The AUV dynamic model will be as a black box that has an input-output relationship instead of using the mathematical model with hydrodynamic parameters. The development of the AUV dynamic model identification will be based upon fuzzy techniques. The second one is designing and developing a robust guidance and control system in order to achieve the desired positions and velocities for the vehicle. Therefore, this paper proposes fuzzy control system. Figure 1 shows a typical underwater vehicle model. One electrical thrusters power the AUV for forward motion. Two electrical pumps used for manoeuvring in the horizontal plane. In addition, two electrical pumps help the AUV to navigate in the vertical plane. The inner box is used for carrying the sensors, battery and the electronic accessories. The hydrodynamic forces per unit mass acting on each axis will be denoted by the uppercase letters X, Y and Z. u, v and w represent the forward, lateral and vertical speeds along x, y and z axes respectively. Similarly, the hydrodynamic moments on AUV will be denoted by L, M and N acting around x, y and z axis respectively. The angular rates will be denoted by p, q and r the components. Dynamics of AUVs, including hydrodynamic parameters uncertainties, are highly nonlinear, coupled and time varying. According to [10], the six degrees-of-freedom nonlinear equations of motion of the vehicle are defined with respect to two coordinate systems as shown in Figure 1. The equations of motion for the UUV are derived from Newton\u2019s second law of motion. The equations of motion for underwater vehicle can be written as follows [11]: (1) where is a 6x6 inertia matrix as a sum of the rigid body inertia matrix, MR and the hydrodynamic virtual inertia (added mass) MA. is a 6x6 Coriolis and centripetal matrix including rigid body terms and terms due to added mass. is a 6x 6 damping matrix including terms due to drag forces. G(q) is a 6x1 vector containing the restoring terms formed by the vehicle\u2019s buoyancy and gravitational terms" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003839_j.snb.2013.11.067-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003839_j.snb.2013.11.067-Figure1-1.png", + "caption": "Fig. 1. Construction of screen-printed electrode.", + "texts": [ + " 78 \u25e6C) was synthesized by Dr. V. Baulin, Instiute of Physiologically Active Compounds RAS, Chernogolovka. Its ynthesis described in [26] as the first step in synthesis of nitrate alt of the same cation. The ISEs with IOS sensing layer (IOS-ISE) ere produced by modification of commercially available SPE, creen-printed electrodes (Elcom, Russia). The SPE have layered esign with silver, graphite and insulator ink layers printed at the olyester substrate; electrode dimensions are 10 \u00d7 28 \u00d7 0.035 mm Fig. 1). Electrode work surface (not more than 0.2 cm2) is free of nsulator. Modification of SPE was performed by putting a piece f solid IL to the electrode\u2019s surface and heating for a few seconds o IL melts and flows over the surface; after removing a heat, IL s quickly solidified forming a layer on the surface. The whole rocedure of preparation of solid-state IOS-ISE is quite simple and akes only 5\u201310 min. For comparison purposes, we also prepared the conventional plasticized membrane ISE (PVC-ISE further) containing the same electrode-active compound, [(C16)2Im]Br" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000009_12.723300-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000009_12.723300-Figure1-1.png", + "caption": "Fig. 1: Geometric illustration of capture", + "texts": [ + " of SPIE Vol. 6578 657811-2 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/20/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx communication structure), compose of communication pairs (p1, p2), (p2, p3), \u2026, (pn-1, pn), will work for formation control. And individual discrepancies are allowed among the pursuers, i.e. they can have different maximal speed. In order to capture a superior evader, a necessary condition was proposed in reference 16 as illustrated by Figure 1. Denote the velocity of the pursuer pv and that of the evader ev . Here, e pv v> . If the pursuer captures the evader as shown in Figure 1, by triangle property, we have sin sin e p v v \u03b1 \u03b2 = , that is, sin sinp p e e v v v v \u03b2 \u03b1= \u2264 (4) Clearly, ( )* max arcsin p ev v\u03b2 \u03b2= = . Thus, it can be seen that if the angle between the moving direction of the evader and the line of sight (LOS) of the pursuer is no larger than *\u03b2 , then the pursuer can always find an angle \u03b1 so that the evader can be captured. To summarize, we must have (5) for the success of capture. arcsin 0 , 2 p p e e v v v v \u03c0\u03b2 \u03b2 \u239b \u239e \u239b \u239e\u2264 \u2264 \u2264 \u2264\u239c \u239f \u239c \u239f \u239d \u23a0\u239d \u23a0 (5) Formula (5) gives a guidance for pursuers\u2019 steering control when capturing a superior evader" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002647_0954406212461326-Figure21-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002647_0954406212461326-Figure21-1.png", + "caption": "Figure 21. The lateral face and end face of 3D crack propagation path.", + "texts": [ + " It is concluded that the type of crack propagation is opening mode mainly in these steps. Thereinto, the changing trend of KI is from big to small firstly, and then gradually increases in every single step. However, with the increase of loading steps, KI is increased relative to the previous step; the whole changing trend of KII basically is the same as KI; the changing trend of KIII is single and degressive, this shows that the direction of crack propagation is stable. It is basically consistent with the experimental results. Figure 21 shows that the crack propagation path of cracked gear under the repeated loads, including the initial crack state and the crack state expanding to the 35th step. When the crack is expended to the 35th step, the crack length reaches a critical value, if the load is still applied at the moment, the gear tooth breaks suddenly. The 3D figures on the left show the lateral expansion of crack surface, and the 2D at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from figures on the right show the expansion of end face of crack at tooth root, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001466_s0019-9958(74)90833-x-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001466_s0019-9958(74)90833-x-Figure3-1.png", + "caption": "FIG. 3.", + "texts": [ + " Suppose for simplicity that each cell has jus t two states, 0 and 1, that a cell in state 1 changes to state 0, and that a cell in state 0 takes on the same state as its left-neighbor, and that the left environment is a 0. Then a filament of such cells, which initially has the leftmost-cell in state 1 and the other cells in state 0 will appear to have a 1-state moving to the right along it at the rate of one cell in unit time. We call this a wave of rate one. I f each cell has more than two states, we can cause waves to move at a different rate through a filament. A slower wave, such as the one represented by line 2 in Fig. 3, moves at a rate of one cell in n units of time. Thus the wave stays in a cell for n t ime units, then moves to the next cell. We can picture the wave as being in the left 1In of a cell, then in position 2In in the cell at the next time instant, and so on up to position n/n, then in the left 1/n of the next cell. The preparatory stage is initiated by a disturbance in the left cell. At the end of the preparatory stage, i.e., at the start of the first-standard stage, markers appear simultaneously at the ends and in the middle of the filament. The method of synchronization of the two end cells with the middle is indicated in Fig. 3. Waves 1, 2 and 3 start at the left end cell at exactly the same time, while waves 4 and 5 start at the right end cell at the time when wave 1 meets the right end cell of the filament. The rates are chosen so that wave 2 meets the right end cell, wave 4 meets the left end cell and waves 3 and 5 meet in the middle of the filament at exactly the same time. Preparatory stage in the synchronization of a growing filament. Before we begin discussing the exact rates of the waves in order to achieve what we want, we have to pay some attention to a problem caused by the discreteness of the cellular array. In Fig. 3, since the waves are represented by straight lines, two waves moving at different speeds will meet at exactly one point. However, Fig. 3 is an idealization of the real situation, just as Fig. 1 is an idealization of Fig. 2. I f waves are propagated by the mechanism indicated in Fig. 2, then it is possible for two waves moving at different speeds to occupy the same cell or neighboring cells for a number of steps. I f we superimpose Fig. 1 on Fig. 2, we see that the boundary waves in Fig. 1 go through the centers of boundary cells in Fig. 2 exactly at those times when the counter in the cell is 0. We simplify our programming by choosing the rates of the waves so that whenever, in a continuous diagram such as Fig. 3, a wave meets an end of the filament, then in the corresponding filament of discrete cells the wave is deemed to meet the end when it reaches the middle of the end cell. Similarly, whenever two waves meet inside a filament they do so in the middle of a cell if a single marker is to be created, or on the edge between two cells if two adjacent markers are needed. In making this simplification, we increase the time needed for synchronization of the whole filament. Let Yi be the speed of propagation of the i-th wave in Fig. 3. (If Yi propagates from right to left, then it shall have a negative value.) We choose the rates of the waves to be l + p Y l - - q q + p x Y 2 - - qx rqx + p t x + tq Ya = 2 tqx ' r x - - p t - - t Y4 - - t ( x - - q) \" and p t x + rqx - - 2pq t - - tq Y~ = 2 tq (x - - q) where x is the least positive integer such that q and t each divide x and ly41 ~ 1 . The reasons for our choice are as follows. Since we wish wave 1 to meet the right end of the filament in the middle of a cell, the time of this event must be a multiple of q, irrespective of the length l of the original filament. Clearly this time is also proportional to l. Our choice of Yl ensures that the time is ql. The time at which wave 2 meets the right end is also proportional to 1. Suppose it is xl. Since x l must be divisible by both q and t for any l, we must have that both q and t divide x. Using x, we can easily work out the values of Y2, Ya, Y~ and y~ from Fig. 3. Since no wave can move through the filament with rate greater than one, we must check that each [Yi I ~ 1. Since p < q and p and q are integers we have l Yl 1 ~ 1. Since p ~ q, q ~ x and p and q are integers we have [Y~ I ~ 1. Since Ya must be less than the maximum of [ r / t ] and 1 Y2 I we have ]Ya ] ~ 1. Our choice of x ensures that ]Y4 [ ~ 1, so it is clear from Fig. 3 that [y~[ ~< 1. By implementing waves with the rates of propagation specified above, 1 1 2 HERMAN ET AL. we can achieve that after xl steps the filament has a marker at the left end, a marker at the right end and one or two markers exactly in the middle, depending on whether the number of cells at the time is odd or even. This is the beginning of the first standard stage. For example, consider Fig. 3. In that figure, r = --1, t = 2, p = 2 and q = 3 . Therefore, Yl = 1, x = 12, Y2 = 3 / 4 , yz = 1/8, Y 4 = - - I and Y5 ----- --1/6. I f the length of the initial filament is 3 cells (l = 2), then the preparatory stage will require 24 steps. At the beginning of the first standard stage there will be 31 cells, three of which are markers, dividing the filament into two equal subsegments, and the length of each subsegment (including the adjacent marker cells) will be 15. Let us assume that we are at the beginning of a standard stage", + " In order to do this we first set up a table of information about the rates at which waves move, and another table which specifies what happens when pairs of waves meet. C E L I A then uses the D E L T A routine to simulate the growth and synchronization of the filament. There are some subtleties in the D E L T A routine. For instance, two waves which are due to meet may occupy the same cell for some time before the actual meeting takes place. Also, in the case where a meeting is to create two markers in adjacent cells, the plan shown in Figs. 3 and 4 calls for two waves to CELIA implementation of a preparatory stage (corresponding to Fig. 3). 643[25/2-2 meet on the edge between two cells. In our D E L T A routine we make the convention that a right-moving wave which would be on an edge in the overall plan, is in fact in the right cell (and a left-moving wave is in the left cell), and we have a special predicate which detects this situation. Figures 6, 7 and 8 show the CELIA print-outs which correspond to the situations described in Figs. 3, 4 and 5, respectively. In the figures we have not, for obvious reasons, instructed CELIA to print all of the attributes of a cell" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000752_detc2007-34090-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000752_detc2007-34090-Figure4-1.png", + "caption": "Figure 4\u2014NASA Glenn Research Center gear fatigue test apparatus. (a) Cutaway view. (b) Schematic view.", + "texts": [ + " The surface was etched in 2% nital. The gear tooth has a uniform lath martensitic structure. Micro-hardness was measured in both the root and flank locations. Vickers hardness measurements were made, and the results were converted to Rockwell C scale. Figure 3 provides the measured hardness as a function of depth from the surface for the flank location. Gear Test Apparatus for Surface Fatigue: The gear surface fatigue tests were performed in the NASA Glenn Research Center\u2019s gear test apparatus. The test rig is shown in Figure 4(a) and is described in Reference [5]. The rig uses the foursquare principle of applying test loads so that the input drive only needs to overcome the frictional losses in the system. The test rig is belt driven and the variable speed motor was operated at a fixed speed for the subject testing. TABLE I\u2014Spur Gear Data. [Gear Tolerance per AGMA 2000-A88 Class 12] Module, mm 3.175 Circular pitch, mm (in.) 9.975 (0.3927) Whole depth, mm (in.) 7.62 (0.300) Addendum, mm (in.) 3.18 (.125) Chordal tooth thickness reference, mm (in", + "asmedigitalcollection.asme.org/ on 01/30/201 Depth from surface (mm) 0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 H ar dn es s (R c) 50 52 54 56 58 60 62 64 66 68 Figure 3\u2014Measured hardness of a FerriumTM C69 test gear tooth. The data are an average value of three measurements made at each reported depth. 2 t and is not subject to copyright protection in the United States. lease; distribution is unlimited. 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use Dow A schematic of the testing apparatus is shown in Figure 4(b). Oil pressure and leakage replacement flow is supplied to the load vanes through a shaft seal. As the oil pressure is increased on the load vanes located inside one of the slave gears, torque is applied to its shaft. This torque is transmitted through the test gears and back to the slave gears. In this way, power is re-circulated and the desired load and corresponding stress level on the test gear teeth may be obtained by adjusting the hydraulic pressure. The two identical test gears may be started under no load, and the load can then be applied gradually. This arrangement has the feature that changes in load do not affect the width or position of the running track on the gear teeth. The gears are tested with the faces offset as shown in Figure 4. By making use of the offset arrangement, the desired contact stress can be achieved within the torque capacity of the testing machine. Because of the offset testing arrangement, four tests can be completed for each pair of gears. Separate lubrication systems are provided for the test and slave gears. The two lubrication systems are separated at the gearbox shafts by lip seals. The two lubrication systems use the same type of oil. The test gear lubricant is filtered through a 5-\u00b5m (200-\u00b5in.) nominal fiberglass filter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002405_s12046-012-0082-4-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002405_s12046-012-0082-4-Figure4-1.png", + "caption": "Figure 4. Schematic of a two-link revolute robot.", + "texts": [ + " Hence equation (36) should be satisfied for all the corner matrices of Anew and Enew for composite system (30) to be asymptotically stable. The matrix Wai is assumed such that constraint (27) is satisfied and the bounding parameter \u03b1i is to be maximized. Hence (36) can be reformulated as an LMI optimization problem as stated in (28). In other words, system (21) is robustly stabilizable by the set of designed decoupled stabilizing PID controllers provided the LMI (28) has a feasible solution for all corner matrices. This completes the proof. Consider a two-link manipulator as shown in figure 4 and its dynamics can be described by nonlinear equation (1). The matrices M(\u03b8), V (\u03b8, \u03b8\u0307) and G(\u03b8) for this two-link robot are M(\u03b8) = [ m11 m12 m12 m22 ] = [ a1 + a2 + 2a3 cos \u03b82 a2 + a3 cos \u03b82 a2 + a3 cos \u03b82 a2 ] , V (\u03b8, \u03b8\u0307) = [ \u2212 (a3 sin \u03b82) ( \u03b8\u03072 2 + 2\u03b8\u03071\u03b8\u03072 ) (a3 sin \u03b82) \u03b8\u03072 1 ] , G(\u03b8) = [ a4 cos \u03b81 + a5 cos (\u03b81 + \u03b82) a5 cos (\u03b81 + \u03b82) ] . (37) In the above expression a1, a2, . . ., a5 are constant parameters obtained from mass (m1, m2) and length (l1, l2) of robot links [ a1 = (m1 + m2) l2 1 , a2 = m2l2 2 , a3 = m2l1l2, a4 = (m1 + m2) l1g, a5 = m2l2g ] " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003535_s10965-013-0083-y-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003535_s10965-013-0083-y-Figure1-1.png", + "caption": "Fig. 1 a XRD patterns of Meso-PANI and Cu/MesoPANI. b TEM micrograph of Cu/Meso-PANI and corresponding diffraction pattern of Cu nanoparticles", + "texts": [ + "3\u00b0 was observed. First peak is related to the periodicity parallel to the polymer chain, while the latter peak caused by the periodicity perpendicular to the polymer chain [7]. Besides reflection corresponding to Meso-PANI [7], several more reflections were also observed in the XRD pattern of Cu/Meso-PANI. The diffraction peaks located at 2\u03b8 values of 43.3\u00b0, 50.4\u00b0, 74.0\u00b0, and 90.0\u00b0 corresponds to (111), (200), (220), and (311) respectively, for metallic Cu in the fcc lattice (JCPDS, File No. 85\u20131326) (Fig. 1a). BET surface area for Cu/Meso-PANI was found to be less than Meso-PANI but much higher than Cu/PANI (Table 1). As shown in TEM images (Fig. 1b), Cu nanoparticles (black spots) are highly dispersed on the Meso-PANI matrix. It can also be seen that the spherically shaped Cu nanoparticles are uniformly dispersed in the Meso-PANI matrix. The particles are not clearly monodispersed. The selected area electron diffraction (SAED) patterns show diffuse ring pattern for the Cu nanoparticles present in Cu/Meso-PANI (Fig. 1b). EDS analysis also confirms that Cu nanoparticles are present on the surface of Cu/Meso-PANI (Fig. 2a). EDS mapping shows the homogeneous distribution of finely dispersed copper nanoparticles in the Meso-PANI matrix (Fig. 2b). The UV-visible spectra of Meso-PANI show two peaks at 356 nm and 444 nm, which are assigned to the \u03c0\u2013\u03c0* benzenoid transition and the benzenoid to quinoid excitotic transition, respectively (Fig. 3a). Besides these two peaks mentioned above, one broad absorption centered at \u03bbmax=738 nm for Meso-PANI was also observed (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002933_s11837-013-0594-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002933_s11837-013-0594-3-Figure1-1.png", + "caption": "Fig. 1. The comparison of coaxial (left) and side (right) laser cladding setup with a moving substrate.5", + "texts": [ + "4 Either the clad material can be introduced before laser processing, i.e., preplaced on the substrate and subsequently melted by the laser beam, or it can be added to the melt pool during the cladding process. The former method has only limited applications as it is time-consuming and difficult to apply for parts with complex geometries.1 In the second method, the clad material is added to the melt pool usually in the form of injected powder particles either by a side nozzle or coaxially with the laser beam as shown in Fig. 1.5 In comparison to conventional surface-deposition technologies such as spray and fuse or arc-welding processes, laser cladding has distinct advantages including minimal dilution from the base metal, lower heat input and hence less distortion and smaller heataffected zone in the substrate component, fully dense coatings with metallurgical bonding to the substrate, excellent control of the layer thickness and composition, better surface quality and tight dimensional tolerances with higher material usage and little or no after-machining, the possibility of selective and precise deposition on sensitive high-value components, etc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002226_s00339-010-5768-z-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002226_s00339-010-5768-z-Figure3-1.png", + "caption": "Fig. 3 3-D schematic representation of the keyhole, and a surface element, with its normal vector", + "texts": [ + " The drilling velocity is proportional to the absorbed laser intensity following the relation: vd = KIabs cos(a) (2) where K is a proportionality factor, Iabs the absorbed laser intensity, and a the beam incidence angle over the sample surface. The local absorbed laser intensity is thus needed in order to calculate the free surface deformation, it is required also the knowledge of the local incident angle \u03b1 of the laser beam over the surface element. We proceed by considering the surface pattern as a 3-D envelope where we need to know the normal vector AN at each surface element as shown in Fig. 3, where Ax , Ay and Az represent the vector components. The modulus of the normal vector AN and also the vector components are data reachable from Fluent, which update the mesh geometry after each time step. Thus one can obtain the incidence angle \u03b1 through the relation: cos(a) = Az/| AN | (3) and one can thus calculate the absorbed intensity such as (for a Gaussian beam): Iabs = A ( 2P/\u03c0r2 1 ) cos1.2(\u03b1) exp (\u22122r2/r2 1 ) (4) Where A is the absorption coefficient of the material, related to the used laser wavelength, P the laser power, and r1 the beam focal spot radius" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002385_ie202394c-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002385_ie202394c-Figure2-1.png", + "caption": "Figure 2. Schematic diagram of reaction apparatus (1, 5: reactants tank; 2, 6: peristaltic pump; 3, 7: valve; 4, 8: flowmeter; 9: RPB; 10: product tank; 11: outlet; 12: wire mesh packing).", + "texts": [ + "75 g of p-PDA at 120 \u00b0C in 150 mL of DMF for 72 h in a nitrogen atmosphere. The resulting products p-MWNTs were filtered and washed with DMF, methanol, 1.0 M HCl solution, and deionized water successively and dried in vacuum (Step c). 2.3. Synthesis of HCl doped PANI Grafted MWNTs Nanocomposites. The nanocomposites were synthesized by in situ oxidation polymerization of aniline in the presence of pMWNTs using RPB as a reactor. The whole experiment process is similar with other literature.15 The reaction apparatus is illustrated in Figure 2. Typically, 1.86 g (0.02M) of aniline and 1.14 g (0.005M) of APS were dissolved in 100 mL of HCl (1.0 M), respectively. Then, various weight ratios of p-MWNTs were dissolved in the aniline solutions and ultrasonicated for 1 h. Circulator bath systems were provided for the whole procedure so as to keep the reaction at 20 \u00b0C. Then, the asprepared solutions were pumped into RPB by two peristaltic pumps (at 100 mL/min each one), premixed by a tee joint, and injected into the RPB (at a high-gravity level of 343 m/s2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.5-1.png", + "caption": "Fig. 2.5 The model of the Verge escapement watch", + "texts": [ + " This also puts the other pallet in position to catch the tooth of the escape wheel on the other side. As the escape wheel continues to rotate, it drives the vertical shaft to rotate in the opposite direction, completing a cycle (Fig. 2.4b). The cycle then repeats converting the rotary motion of the escape wheel to the oscillating motion of the verge. Each cycle advances the wheel train of the clock moving the hands forward at a constant rate. The Verge escapement was first used as a clock escapement and then modified into a watch escapement. Figure 2.5 shows the Verge watch escapement. From the figure, it can be seen that the crown-shaped escape wheel and the vertical shaft are the same; but the horizontal bar is replaced by a balance wheel with a hairspring. In this case, the timekeeping is regulated in part by the hairspring as it controls the engagement of the second pallet. A computer animation is shown on the Springer Website http://extra.springer.com/2012/978-3-642-29307-8. Figure 2.6 shows a Verge escapement clock made in late 1700s (Institute of Precision Engineering 2008)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002277_icit.2012.6210043-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002277_icit.2012.6210043-Figure3-1.png", + "caption": "Fig. 3. Rotational Motion Principles", + "texts": [ + " Special attention must be paid to the following concept. The tilting of the tail rotor is not utilized in the yawing motion control as in typical Tri-Rotor designs, but in the decoupling of the rolling motion from the pitching tail thrust. This is achieved by controlling the tail rotor's rotation angle via the tail servo so as to track the rolling angle and produce an opposite and equal rotation angle. Through this, we ensure that the tail rotor thrust is always applied as a pitching motion component. The aforementioned concepts are presented in Figure 3. Let B = {Bx, By, Bz} be the coordination body-fixed reference frame and E = {Ex, Ey, Ez} be the Flat-Earth model inertial reference frame as depicted in Figure 4. It should be noted that the earth inertial frame (EFF) follows the North East-Down (NED) notation and the body-fixed inertial frame (BFF) follows the standard aircraft notation where the z axis points downwards, the x-axis towards the longitudinal flight direction and the y-axis towards the right wing. Also let U = {u, v, w} be the vector of linear rates and Q = {p, q, r} the vector of angular rates expressed on the coordination frame B and let X = {x, y, z} be the vector of translational displacements and e = {/{), e, lJI} the vector of rotational displacements expressed on the inertial frame E" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001935_1.3478635-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001935_1.3478635-Figure1-1.png", + "caption": "Figure 1. Color online Schematic of electrochemical cell used for impedance measurements, illustrating the positions of the two working electrodes WE1 and WE2 , the reference electrode RE , and the counter electrode CE .", + "texts": [ + " The electrode was then immersed for 1 h into a solution containing 50 g/mL antibody and 50 mM PBS at pH 7.3, which immobilizes antibodies onto the Au electrode by amide bond formation. This was then immersed into a solution containing 14.7 g/mL BSA to block unreacted sites on the Au electrode. These antibody-coated Au electrodes were then exposed to peanut protein Ara h 1 in filtered and diluted soup. Measurement methods.\u2014 All electrochemical measurements were performed with a three-electrode configuration using a Pt spiral counter electrode and a saturated calomel reference electrode. Figure 1 illustrates the electrochemical cell that was constructed with two working electrodes, with one working electrode serving as the measurement channel and the other serving as a reference channel. The counter electrode is 2 to 3 mm from one working electrode WE measurement channel and 4 to 6 mm from the other working electrode reference channel . Impedance measurements were performed using an EG&G PAR 273A potentiostat coupled to a Solartron 1250 frequency response analyzer at a constant frequency of 5 Hz. Impedance measurements were multiplexed between the two working electrodes in Fig. 1 using the ZPlot software. Whenever the working electrode was switched, the system was allowed to stabilize before impedance measurements were recorded. Test solution.\u2014 The test solution was obtained from a can of chicken soup purchased at the supermarket. The soup was filtered successively through a kitchen strainer, nylon mesh filter with a pore size of 20 m, and a nanoporous alumina filter with a nominal pore ) unless CC License in place (see abstract).\u00a0 ecsdl.org/site/terms_uses of use (see Downlo size of 100 nm", + "23-26 This is accomplished by immobilizing a closely similar antibody in the reference channel whose antigen is absent from the samples of interest and subtracting the reference channel from the measurement channel.23-26 In addition to correcting for the effects of nonspecific adsorption, this method also compensates for variations in temperature and refractive index. SPR spectroscopy can detect changes in the optical properties of an antibody film upon antigen binding, whereas EIS can detect changes in the electrical properties of an antibody film. This analogy motivated the use of the electrochemical cell shown in Fig. 1, where address. Redistribution subject to ECS term130.237.29.138aded on 2015-03-07 to IP the antibody to peanut protein Ara h 1 is immobilized at one Au working electrode measurement channel , and the antibody to cockroach protein Bla g 1 is immobilized at the other Au working electrode reference channel . In addition to correcting for the effects of nonspecific adsorption, the reference channel in Fig. 1 may also help correct for variations in the sample viscosity and ionic strength. Impedance results at Au electrodes with immobilized antibodies.\u2014 Upon addition of peanut protein Ara h 1 to the test solution, the real component of the impedance at 5 Hz increases both for the measurement channel, at which the antibody to peanut protein Ara h 1 is immobilized, and for the reference channel, at which the antibody to cockroach protein Bla g 1 is immobilized. The cumulative impedance increase at 5 Hz upon introduction of peanut protein Ara h 1 at the two different working electrodes is quantitatively compared in Tables I and II for the three different dilution levels" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003299_lindi.2012.6319476-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003299_lindi.2012.6319476-Figure1-1.png", + "caption": "Fig. 1. The robotic car we used for testing purposes", + "texts": [ + " potential field and wave propagation) methods and fuzzy methods are used as well. At the start of the article, we shortly describe the recent publications and results in this topic, and then we introduce a fuzzy-based navigational system. In the second part of the article, we compare an earlier fuzzy system with our improved version. The mentioned algorithms were tested in the MATLAB environment, the improved version was tested in a real-world environment as well using a 4-wheeled autonomous robotic car (see Figure 1.) To implement an efficient path-planning algorithm, first we have to examine the types of the different maps that we can build, as we can only select the appropriate algorithm if we consider the map type as well. According to sources [1] and [2], we should use particle-based maps when navigating on smaller areas, and we should use the graph/landmarkbased maps if we have a bigger map area. Due to some data correlation problems, the graph-based methods can only be used in an offline way, so the methods described in this article use particle-based, binary occupancy grid maps" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003589_1077546311405701-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003589_1077546311405701-Figure1-1.png", + "caption": "Figure 1. A rotor-ball bearing system.", + "texts": [ + " The key point is to analyze the influence of the nonlinear ball bearing force and Alford force in the paper. As a result, it makes an assumption that the torsional vibration of rotor and gyroscopic effects may be neglected and only the transverse vibration of rotor should be considered. Thus the rotor system can be modeled as a Jeffcott rotor system, in which the rotor is simplified to one disk with two transverse stiffness and the masses of the shaft being equivalent to the disk and two bearings. The model of rotor system is shown in Figure 1. The two ends of this rotor are supported on similar ball bearings. The mathematical model of the rotor system takes into account 4 degrees-of-freedom \u2014horizontal and vertical displacements of the rotor at the disk location (X1,Y1) and at the ball bearing (X2,Y2), correspondingly. After assembling the Alford force and ball bearing force, the dynamic equations of the system are established as follows: M1 \u20acX1 \u00fe C1 _X1 \u00fe 2K X1 X2\u00f0 \u00de \u00bcM1e! 2 cos !t\u00f0 \u00de M1 \u20acY1 \u00fe C1 _Y1 \u00fe 2K Y1 Y2\u00f0 \u00de \u00bcM1e! 2 sin !t\u00f0 \u00de M1g\u00fe FaY1 M2 \u20acX2 \u00fe C2 _X2 \u00fe K X2 X1\u00f0 \u00de \u00bc FbX2 M2 \u20acY2 \u00fe C2 \u20acY2 \u00fe K Y2 Y1\u00f0 \u00de \u00bc FbY2 M2g 9>>>= >>>; : \u00f01\u00de The ball bearing model considered here has equispaced balls rolling on the surfaces of the inner and outer races" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003029_978-3-642-33959-2_7-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003029_978-3-642-33959-2_7-Figure1-1.png", + "caption": "Fig. 1 Quadrotor rotorcraft \u2013 a non-linear dynamic system that uses synergy of its rotarywings to fly. Commonly used model of quadrotor with corresponding degrees of freedom. Coordinate systems assumed to enable model derivation.", + "texts": [ + " The usage of the computed maneuvers as a benchmark is demonstrated by evaluating quadrotor design parameters, and a linear feedback control law as an example of a control strategy. The main contribution of the paper regards to design and simulation of the appropriate simulation benchmark procedures (indoor as well outdoor) to be used for the objective assessment and evaluation of different control algorithms applied to microcopoter rotorcrafts. The quadrotor is satisfactory well modeled [21] - [23] with a four rotors in a cross configuration Fig. 1. This cross structure is quite thin and light, however it shows robustness by linking mechanically the motors (which are heavier than the structure). Each propeller is connected to the motor through the reduction gears. All the propellers axes of rotation are fixed and parallel. Furthermore, they have fixed-pitch blades and their air flows point downwards (to get an upward lift). These considerations point out that the structure is quite rigid and the only things that can vary are the propeller speeds. As shown in Fig. 1, one pair of opposite propellers of quadrotor rotates clockwise (2 and 4), whereas the other pair rotates anticlockwise (1 and 3). This way it is able to avoid the yaw drift due to reactive torques. This configuration also offers the advantage of lateral motion without changing the pitch of the propeller blades. Fixed pitch simplifies rotor mechanics and reduces the gyroscopic effects. Control of quadrotor is achieved by commanding different speeds to different propellers, which in turn produces differential aerodynamic forces and moments", + " In order to pitch and move laterally in that direction, speed of propellers 3 and 1 is changed conversely. Similarly, for roll and corresponding lateral motion, speed of propellers 2 and 4 is changed conversely. To produce yaw, the speed of one pair of two oppositely placed propellers is increased while the speed of the other pair is decreased by the same amount. This way, overall thrust produced is same, but differential drag moment creates yawing motion. In spite of four actuators, the quadrotor is still an under-actuated system. The Fig. 1 shows the structure model [22] [23] in hovering condition, where all the propellers have the same speed of rotation \u03c9i = \u03c9H , i = 1, . . . ,4. In the Fig. 1 all the propellers rotate at the same (hovering) speed \u03c9H (rad/s) to counterbalance the acceleration due to gravity. Thus, the quadrotor performs stationary flight and no forces or torques moves it from its position. Even though, the quadrotor has 6 DOFs, it is equipped just with four propellers hence it is not possible to reach a desired setpoint for all the DOFs, but at maximum four. However, thanks to its structure, it is quite easy to choose the four best controllable variables and to decouple them to make control easier. The four quadrotor targets are thus related to the four basic movements which allow the microcopter to reach a certain height and attitude. To describe the motion of a 6 DOF rigid body it is usual to define two reference frames Fig. 1 : (i) the earth inertial frame (E-frame), and (ii) the body-fixed frame (B-frame). The linear position of the helicopter (X ,Y,Z) is determined by the coordinates of the vector between the origin of the B-frame and the origin of the E-frame according to equation. The angular position (or attitude) of the helicopter ( \u03d5 , \u03b8 , \u03c8) is defined by the orientation of the B-frame with respect to the E-frame. The vector that describes quadrotor position and orientation is: s = [X Y Z \u03d5 \u03b8 \u03c8 ]T . (1) The generalized quadrotor velocity expressed in the B-frame can be written as [1]: v = [u \u03c5 w p q r]T , (2) where, the u, \u03c5 , w represent linear velocity components in the B-frame, while p, q, r are corresponding angular velocities of rotation about corresponding roll, pitch and yaw axes", + " A proportional-integrative-derivative controller (PID) is most common feedback form in all kinds of control systems, and is also being used for flight control of quadrotor [4] [5]. The ideal PID is represented in continuous time domain as: u(t) = kPe(t)+ kI t\u222b 0 e(\u03c4)d\u03c4+ kD de(t) dt , (10) where the terms kP, kI and kD are proportional, integral and derivative gain, respectively, u(t) is the output of the controller, and input e(t), is the reference tracking error. The PID quadrotor flight controller consists of six PID controllers for the particular state coordinates (1), see Fig. 1. The backstepping technique is recursive design methodology that makes use of Lyapunov stability theory to force the system to follow a desired trajectory. Backstepping approach to quadrotor flight control was successfully applied in number of researches such as for example [10] [24]. First, the dynamical model from (5) is rewritten in state-space form X\u0307 = f (X ,U), by introducing X = [x1, . . . ,x12] T \u2208\u211c12as space vector of the system (7) [24]: x1 = \u03c6 , x2 = x\u03071 = \u03c6\u0307 , x3 = \u03b8 , x4 = x\u03073 = \u03b8\u0307 , x5 = \u03c8 , x6 = x\u03075 = \u03c8\u0307 , x7 = X , x8 = x\u03077 = X\u0307 , x9 = Y, x10 = x\u03079 = Y\u0307 , x11 = Z, x12 = x\u030711 = Z\u0307" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003945_icra.2013.6631143-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003945_icra.2013.6631143-Figure4-1.png", + "caption": "Fig. 4. Section frames", + "texts": [ + " If segi has a non-zero curvature, we call the circle of segi the section i\u2019s circle, denoted by ciri, with radius ri, and the plane that contains segi the section plane, denoted by Pi, as shown in Fig. 3. The base frame of the robot is set at p0 with z0 axis tangent to seg1. The frame of section seci, i = 2, 3, is formed at pi\u22121 with the zi axis tangent to segi at pi\u22121. The base of segi is the tip of segi\u22121. Adjacent segi\u22121 and segi are connected tangentially at the connection point pi\u22121 as shown in Fig. 4(a), i.e., the two sections share the same tangent at pi\u22121. For clarity, we consistently use black, red, and green colors to draw section 1, section 2, and section 3 of the OctArm in this paper. Each section i has three degrees of freedom that can be directly changed by the OctArm actuators [8], which are controllable variables: curvature \u03bai, length si, and orientation angle \u03c6i from yi\u22121 axis to yi axis about zi axis. Fig. 4(b) shows one example segi, its frame, and controllable variables. Note that the center ci of ciri always lies on the xi axis, with \u22131/\u03bai being the x coordinate in the i-th frame, where \u03bai is the curvature. Note also that ci lies on the positive xi axis if \u03bai < 0 and on the negative xi axis if \u03bai > 0. When \u03bai = 0, segi is a straight-line segment along the z axis, and ci can be considered at either +\u221e or \u2212\u221e along the x axis. The configuration of the entire arm is determined by the control variables of each section" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003160_bi00826a019-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003160_bi00826a019-Figure3-1.png", + "caption": "FIGURE 3: Variation of the cycle time with the reciprocal value of NADH oxidase activity after inhibition with rotenone. Conditions were as in Figure 2 but a different preparation of electron-transport particle was used: A, calculated for specific inhibition; B, calculated for specific and unspecific inhibition (i.e., no bovine serum albumin wash).", + "texts": [ + " The hyperbolic relation between cycle time and oxidase activity expected from this line of reasoning is borne out in experiments with specifically bound rotenone (Figure 2). When both specifically and unspecifically bound rotenone were present the cycle time at all rotenone concentrations 4752 B I O C H E M I S T R Y , VOL. 9, N O . 2 4 , 1 9 7 0 R E D O X C Y C L E O F N A D H D E H Y D R O G E N A S E tested was very much longer. In this case the hyperbolic relationship is also no longer evident. This may be better seen in Figure 3, which is a plot of reciprocal oxidase activity L.S. cycle time at a series of rotenone concentrations. Curve A, representing specifically bound rotenone, is linear but B, representing specifically and unspecifically bound inhibitor (i.e., bovine serum albumin was omitted) is not. The inferences drawn from these experiments are, first, that most of the lag in the reoxidation of the chromophore is due to unspecifically bound rotenone, second, that when only specifically bound inhibitor is present the lag time represents a steady-state reduction of the enzyme and its duration is determined solely by the rate of electron flux to the respiratory chain, and, third, that unspecifically bound rotenone has an additional, more complex effect on the redox cycle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002727_1.4023084-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002727_1.4023084-Figure1-1.png", + "caption": "Fig. 1 Contact vibration model of a rolling element and the raceway", + "texts": [ + " The change of mutual approach is selected as the standard of vibrations and the inlet length and dimensionless natural frequency corresponding to the working load and speed are determined. The DC-FFT method is implemented to increase the computational efficiency associated with elastic deformations and the semisystem approach is applied to improve solution convergence, which make the analysis of stiffness and damping in wider ranges of load and speed possible. The problem to be considered in this paper is the contact vibration between a rolling element and the raceway, as depicted in Fig. 1. The contact pair is under elastohydrodynamic lubrication and the film can be replaced by a spring and a damper, as shown in Figs. 2(a) and 2(b). In point of the mutual approach, it is below zero when the elastic deformation at the contact center is smaller than the central film thickness; otherwise it is above zero. For the oil film force, Ffilm \u00bc \u00f0\u00f0 X pdxdy the Taylor series is as follows: Ffilm \u00bc FS \u00fe @F @h0 h0 h0S\u00f0 \u00de \u00fe @F @ _h0 _h0 \u00fe 1 2 @2F @h2 0 h0 h0S\u00f0 \u00de2 \u00fe 1 2 @2F @ _h2 0 _h2 0 \u00fe @2F @h0@ _h0 h0 h0S\u00f0 \u00de _h0 \u00fe (1) In small oscillations which often occur in mechanical systems, the oil film force can be linearized around the equilibrium position and the two-order and higher-order terms in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002591_10402004.2011.626144-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002591_10402004.2011.626144-Figure11-1.png", + "caption": "Fig. 11\u2014Effect of shear thinning and viscoelastic properties on journal center trajectory.", + "texts": [ + " 10 it can be observed that the difference between the size of journal center whirl orbits of smooth and rough bearings (moving roughness with transverse D ow nl oa de d by [ N ew Y or k U ni ve rs ity ] at 1 6: 43 0 5 O ct ob er 2 01 4 roughness pattern) is large when surface roughness and nonNewtonian behavior of the lubricant effects are considered together, whereas it is comparatively smaller when the surface roughness effect alone is considered. These results clearly indicate an interactive influence of the surface roughness and nonNewtonian behavior of the lubricant on the bearing response. Therefore, the bearing designer needs to study the influence of surface roughness and non-Newtonian behavior of the lubricant together. Figure 11 indicates that when the value of the power law index, n, is reduced, the size of the journal center whirl orbit increases for both viscoelastic and nonviscoelastic lubricant. For the given power law index of the lubricant, the journal center is observed to whirl in the location of the bearing angle between 180 and 360\u25e6 (i.e., in the third and fourth quadrants of the bearing) when the viscoelastic properties of the lubricant are not considered. When the viscoelastic properties are considered, the journal center is observed to whirl in the location of the bearing angle between 180 and 270\u25e6 (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001612_jsts.25.1_27-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001612_jsts.25.1_27-Figure4-1.png", + "caption": "Fig. 4. ETS-VIII Configuration", + "texts": [ + " Consequently, less-conservative design and the frequency-shaping technique can be incorporated. As a result of our proposed method, we can improve the control performance easily. Moreover, the controller that we design as a L TI controller is more computationally efficient than the LPV controller. Simultaneously, we study the design method that incorporates model change based on differences of mass properties between Beginning Of mission Life (BOL) and End Of mission life (EOL). We can therefore guarantee robust stability of the system for whole its mission life. 2. SPACECRAFT MODEL . Figure 4 depicts the ETS-VIII configuration. It has two solar array paddles and two large deployable antenna reflectors. The solar paddles are deployed in the pitch direction and rotate 360 degrees per day so that each paddle constantly faces the Sun. In general, the equation of motion for the flexible spacecraft is described as the following hybrid equation incorporating rigid-body rotation and the vibration equation of flexible components (solar paddles and antenna reflectors) by the FEM numerical analysis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001124_1.45041-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001124_1.45041-Figure1-1.png", + "caption": "Fig. 1 Transfer orbit geometry of the Lambert problem.", + "texts": [ + " Given two fixed points in space, themultiple-revolution transfer time can be expressed as a function of the transverse-eccentricity component eT and the number of admissible revolutions N, and then the corresponding derivative of the transfer time relative to the transverse eccentricity is derived. Moreover, the transverseeccentricity-based numerical procedure to obtain the solutions of the multiple-revolution Lambert\u2019s problem is proposed, and numerical simulations are carried out in the end to validate it. Figure 1 shows the basic orbital geometry of the classical Lambert problem. In Fig. 1,F is the center of the gravitation,P1 andP2 are the given initial andfinal positions in space, c is the chord length between the two points, and the radius vectors r1 and r2 locate the points P1 and P2 with respect to the focus F. The angle between radius vectors is the transfer angle, and e is the eccentricity vector of the orbit. As described in [1], the component of the eccentricity vector projecting onto the chord P1P2 is constant and can be expressed as eF e ic e r2 r1 =c r1 r2 =c (1) where ic is the chord unit vector, c kr2 r1k, and r1 and r2 are the magnitudes of the corresponding radius vectors. The component of the eccentricity vector perpendicular to the chord is defined as the transverse eccentricity eT. Thus, the eccentricity vector of the transfer orbit passing through P1 and P2 can be decomposed into a constant component parallel to the chord eF and a variable transverse component perpendicular to the chord eT , then the vector can be written as e eFic eTip (2) where ip is the unit vector lying in the orbit plane and perpendicular to the chord direction, as shown in Fig. 1. One gets the minimum eccentricity admissible transfer orbit when the transverse eccentricity is equal to zero. Battin [3] defined this minimum eccentricity orbit as the fundamental ellipse, for which the eccentricity, semimajor axis, and semilatus rectum are given by8< : ef r1 r2 =c af r1 r2 =2 pf af 1 e2f (3) where the subscript f denotes the fundamental ellipse solution. As discussed in [1], the semilatus rectum of a generic transfer orbit can be expressed as a function of the transverse eccentricity as follows: p eT pf eT r1r2 c sin (4) The semimajor axis and orbit period can be expressed as a eT p eT = 1 e2F e2T T eT 2 a3= p (5) where is the gravitational constant", + " Student Member AIAA. \u2020Ph.D. Candidate, School of Astronautics; lijian@sa.buaa.edu.cn. \u2021Professor, School of Astronautics; hanchao@buaa.edu.cn. Vol. 33, No. 1, January\u2013February 2010 265 D ow nl oa de d by B R IS T O L U N IV E R SI T Y o n M ar ch 1 , 2 01 5 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /1 .4 50 41 The true anomalies of P1 and P2 are given by 1 tan 1 eF sin!c eT cos!c;eF cos!c eT sin!c 2 1 (6) where !c is the angle between the radius vector r1 and the chord unit vector ic, as shown in Fig. 1, and tan 1 y; x is the four-quadrant inverse tangent function. Then the exocentric anomaly can be calculated by the relationship between true and exocentric anomaly as follows: E 2tan 1 1 e 1 e r tan 2 (7) From Eqs. (1\u20137), one can see that given the radius vectors r1 and r2, the exocentric anomalies E1 and E2, the semimajor axis, and the eccentricity of the transfer orbit can be expressed as a function of the transverse eccentricity eT. With the help of E1, E2, a, and e, it is convenient to obtain the direct-transfer time ts12 between P1 and P2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002051_cdc.2011.6161047-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002051_cdc.2011.6161047-Figure2-1.png", + "caption": "Fig. 2: Examples of setups discussed by Theorem 3.10.", + "texts": [ + " The proof for the second part follows from the fact that when the control and observation nodes are connected to the same nodes, we have r = q. This results in a quadratic form rTQ\u22121q = rTQ\u22121r, and since Q\u22121 is positive definite, the quadratic form is also positive definite. Note that the vector r cannot be zero, as we are assuming a connected graph G, and r = q = 0 would imply that the control and observation nodes are disconnected from the rest of the graph. Therefore rTQ\u22121r >0, contradicting both conditions (17) and (18). Figure 2 shows examples for the two setup categories discussed by Theorem 3.10. This result can be applied to certain graphs. Corollary 3.11: The system (1) when the underlying graph is the complete graph or the star graph (Figure 1) is minimum phase for arbitrary transmission pair. Note that this corollary is directly verified in Table I from the analytic expressions of the transfer functions. Observe that the star graph corresponds to the first case in Theorem 3.10, and the complete graph to the second" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000819_aim.2009.5229935-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000819_aim.2009.5229935-Figure1-1.png", + "caption": "Fig. 1. Components of harmonic drive gearing.", + "texts": [ + ". INTRODUCTION Harmonic drive gearings are widely used in a variety of industrial applications, e.g. industrial and/or humanoid robots, precision positioning devices, etc., because of their unique kinematics and high performance attributes such as simple and compact mechanism, high gear ratios, high torque, and zero backlash. As shown schematically in Fig.1, this transmission system is generally comprised of just three components: a wave generator (WG) with an elliptical shape, a flexspline (FS) of an elastic thin-walled steel cup, and a circular spline (CS) of a rigid steel ring with internal teeth, which was developed to take advantages of the elastic dynamics of metal. The basic principles of motion are as follows: FS is deflected by WG into an elliptical shape and the inscribed ellipse contacts with internal teeth of CS at two points. Then, the tooth engagement position moves by turns relative to CS, while FS moves by two teeth relative to CS because of two fewer teeth of FS than that of CS" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.39-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.39-1.png", + "caption": "Fig. 2.39 The model of the Dual Ulysse escapement", + "texts": [ + " Oechslin received his Ph.D. in 1983 and his master watchmaker title in the subsequent year. Presently, he is the curator of the Mus\u00e9e International d\u2019Horlogerie, in La Chauxde-Fonds, Switzerland. The dual Ulysse escapement is perhaps inspired by the independent double wheel escapement invented in 1800s. Figure 2.38 shows the model of the double wheel escapement. Like many old designs, the independent double wheel escapement was abandoned because of its complexity and lack of reliability. As shown in Fig. 2.39, the dual Ulysse escapement consists of a balance wheel with a plate and a hairspring, a triangle-shape lever with two horns and two recesses and two escape wheels. There are also two pins used to limit the swing of the lever. Its most notable feature is the two escape wheels with specially designed tooth profile. Escape wheel 1 is driven by the gear train and meshes with Escape wheel 2. The two escape wheels also interlock with each other under the control of the lever. The lever receives pulses generated alternately by the first and the second escape wheels and transmits these pulses to the plate on the balance wheel, driving the balance wheel to swing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002924_mcs.2012.2234971-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002924_mcs.2012.2234971-Figure1-1.png", + "caption": "Figure 1 The coordinate system of the quadrotor platform. Motor numbering, definition of axes, and the axis directions are completed after the inertial sensor is mounted on the body. Arrow tips indicate the positive direction.", + "texts": [ + " Pedagogical issues are addressed in the final section, such as the learning objectives/outcomes and student feedback analysis. Additionally, see \u201cThe Educational Environment,\u201d which describes the physical appearance and user interface of the setup. dESCRIPTION Of ThE EduCATIONAL SETuP The dynamics of a real flying quadrotor has 6 degreesof-freedom (DOF) movement of a rigid body in space [24]. The actual setup is fixed to a mount, so only 3 DOF movement is available. The vertical thrust generated by four rotors, and additional disturbing aerodynamic forces in Figure 1, act on the rigid body. The setup is fixed on a 3-DOF universal joint with !12\u00b0 freedom in the roll and pitch axes, and with !360\u00b0 freedom in the yaw axis. Actually, the yaw axis is limited by only the length of the cables attached to the fixture. Movement is not available on the z-axis. This limited configuration is ideal for attitude control since the dynamic equations are reduced by the elimination of unused axes that are x, y, and z [24]. The roll, pitch, and yaw movements are accomplished by varying the rotor speeds", + " \u00bb The body is fixed on a universal joint, and no translational motion is available. \u00bb Blade flapping and friction are neglected. \u00bb The vertical thrust and angular moment in the hori- zontal plane of each engine are proportional to the square of the rotor angular speed. \u00bb Effects of the ground on the setup are ignored. The propellers are assumed to be high above the ground. \u00bb The roll and pitch angles are assumed to be smaller than 10\u00b0. The body and inertial coordinate systems for the computer-aided design (CAD) image of the setup is shown in Figure 1. The arrow tips indicate the positive direction. This convention matches the inertial sensor\u2019s directions used in the setup. The complete derivation and simplification of the dynamic model can be found in [26]. The same equations are also used in [20], [24], [25], and [27]. The equations of motion are ( ) ( ) I I I I I I bl .. xx rotor . xx yy zz . xx 1 3 2 4 2 2 4 2 z i }i X X X X X X = + - - + - + - o (1) ( ) ( ) I I I I I I bl .. yy rotor . yy zz xx . yy 1 3 2 4 3 2 1 2 i z }z X X X X X X = - - + + + - + - o (2) ( ) ,I d I I I zz zz xx yy ", + " Quadrotor dSpace ds1103 PC Running Matlab and ControlDesk PC RunningEPOS 68 IEEE CONTROL SYSTEMS MAGAZINE \u00bb April 2013 thrust coefficient, d is the drag coefficient, and l is the arm length. The b and d coefficients depend on many variables such as temperature, air density, and propeller characteristics [28]. The body inertia values are denoted by Irotor, Ixx, Iyy, and Izz. Due to the symmetrical properties of the setup, the inertia matrix is diagonal, where Ixx and Iyy are assumed to be equal [26]. The forces and moments acting on the body in Figure 1 can be written as F bi i 2X= (4) di i 2x X= (5) ( ) ( )l F F bl2 4 2 4 2 2x X X= - = -z (6) ( ) ( )l F F bl3 1 3 2 1 2x X X= - = -i (7) di i 1 2 3 4 1 2 2 2 3 2 4 2 1 4 x x x x x x X X X X= = - + - = - + -} = ^ h/ (8) where Fi and ix represent the vertical thrust and horizontal moments produced by each rotor, respectively. The moments in the orientation angles are represented by , ,x xz i and .x} The equations of motion (1)\u2013(3) include structure-specific parameters. The value for l was measured with a ruler, while b and d were calculated using (4)\u2013(5) with the aid of sensor data", + " However, the mechanical part of the setup allows 12\u00b0 freedom, which is approximately 6\u00b0 more than the amplitude of the reference signals. as the system moves to higher orientation angles above 6\u00b0, the closed-loop performance of the proportional-derivative (PD) controller drops. The amplitude of the reference signals was limited to 1 rad (5.7\u00b0) since demonstration of an optimal PD performance was desired. High orientation angles are better controlled with nonlinear control methods, which is beyond the scope of an undergraduate education. No. excited axes reference Signal graphical representation Figure 1 roll Step roll for Kp : 10, 12, 14 Figure 10 2 Yaw Step Yaw for Kp : 0.2, 0.4, 0.6 3 roll Step roll, pitch, yaw for Kp : 12 Figure 11 4 Yaw Step roll, pitch, yaw for Kp : 0.4 5 roll Sine (0.75 Hz) roll for Kp : 10, 12, 14 Figure 12 6 Yaw Sine (0.25 Hz) Yaw for Kp : 0.2, 0.4, 0.6 7 roll pitch Yaw roll: 0.75 Hz pitch: 0.75 Hz, 45\u00b0 Yaw: 0.25 Hz roll + Uroll pitch + Upitch yaw + Uyaw rotor speeds Figure 13 76 IEEE CONTROL SYSTEMS MAGAZINE \u00bb April 2013 centered on the y-axis for better visualization, whereas no scaling is done on the time axis", + " Then the inertial sensor is powered off, and it is shown that the setup can still be controlled but is almost impossible to stabilize. After the theory, the setup is run, and the control structure in Figure 7 and the Simulink model in Figure 9 are projected on the board. The closed- and open-loop control concept is shown on the model and is applied on the setup at the same time. A discussion about open- and closedloop control systems in daily applications is elicited. Topic: Dynamic Modeling Before the theory, the setup is run, and the quadrotor coordinate system in Figure 1 and (1)\u2013(3) are projected on the board. The symbols are briefly explained, and then the students are asked to brainstorm about the dynamic model while performing roll, pitch, and yaw maneuvers. After the theory, the effect of each parameter in (1)\u2013(3) is demonstrated. The goal is for the students make an analogy between the dynamic model and system behavior. First, the students are asked how the system response changes when a specific parameter or term in (1)\u2013(3) is increased or decreased. Then the parameter or term is changed in real time from the ControlDesk interface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000366_elan.200804278-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000366_elan.200804278-Figure2-1.png", + "caption": "Fig. 2. Cyclic voltammograms of a CuHCF film synthesized from 0.1 M aqueous KNO3 solution on the carbon paste electrode, which was then transferred to 0.1 M aqueous solution of a) NaNO3, b) KNO3, c) NH4NO3, and d) Mg(NO3)2, v\u00bc 50 mV s 1.", + "texts": [ + " All the single redox couples are attributed to the [Cu(II)-CN-Fe(III)]/Cu(II)-CNFe(II)] reaction [33 \u2013 35]. The results show the successful formation of CuHCF films and the voltammetric properties Electroanalysis 20, 2008, No. 18, 1996 \u2013 2002 www.electroanalysis.wiley-vch.de D 2008 WILEY-VCH Verlag GmbH & Co. KGaA, Weinheim of film depend on the kind of electrolyte in aqueous solution (Table 1). The electrochemical responses of the CuHCF/CPE in an electrolyte solution with different cations are shown in Figure 2. It was found that one pair of well-defined redox peaks can be observed from Figure 2, curve b, when a K\u00fecontaining electrolyte solution was used. When NH4 \u00fe (Fig. 2, curve c) or Mg2\u00fe (Fig. 2, curve d) containing electrolyte solution was used, one pair of broad and illdefined redox peaks were observed and when electrolyte solution include Na\u00fe (Fig. 2, curve a), two pairs of broad redox peaks were observed. Furthermore, the peak currents decreased drastically compared with that in K\u00fe containing electrolyte solution. The above results indicated that the size of K\u00fe matches best with the cavity of the CuHCF film. The cyclic voltammograms of the CuHCF film synthesized from 0.1 M KNO3 on a carbon paste electrode and then transferred to 0.1 M aqueous KNO3, KCl, K2SO4, KClO4 and KBrO3 solutions are compared. The cyclic voltammograms almost overlap at the redox couple near to 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure9-1.png", + "caption": "Fig. 9 Metamorphic mechanism with two configurations", + "texts": [ + " 8 a , the revolute pair between Links 1 and 5 is frozen. The mechanism is termed as Configuration 5 and is shown in Fig. 8 b . If this configuration comes from Configuration 4, the adjacency matrix A5 can be obtained by multiplying \u22121 on A4 4,5 , A4 5,4 , A4 1,5 , and A4 5,1 , i.e., A5 = 0 1 0 0 \u2212 1 1 0 1 0 0 0 1 0 1 0 0 0 1 0 1 \u2212 1 0 0 1 0 12 Note that among all possible configurations, any one can be transformed to any other one. Though, the adjacency matrix can always catch the process of the change. Figure 9 shows such an example. This is a planar metamorphic mechanism with five links. There is a spring embedded in Link 1, which can push Slider 2 moving along the slot. There are two pins, P1 and P2, on the frame to limit the swing of Link 4. When Link 4 is blocked by one of the pins P1 or P2, Slider 2 will slide along the slot of Link 1, as shown in Figs. 10 and 11. The mechanism works as a guide-bar mechanism in these cases. Here, only two configurations are displayed: If Link 4 does not touch both pins P1 and P2, Slider 2 will rest at the end of the slot JULY 2008, Vol" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000432_s11071-008-9400-0-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000432_s11071-008-9400-0-Figure2-1.png", + "caption": "Fig. 2 Model of a flexible rotor supported by two turbulent journal bearings", + "texts": [ + " Therefore, fe = \u2212\u03bcL3R 2c2 \u222b \u03c0 0 { [(\u03c9 \u2212 2\u03d5\u0307)\u03b5 sin \u03b8 \u2212 2\u03b5\u0307 cos \u03b8 ] cos \u03b8 Gz[(1 + \u03b5 cos \u03b8)3 \u2212 12(l\u2217)2(1 + \u03b5 cos \u03b8) + 24(l\u2217)3 tanh( 1+\u03b5 cos \u03b8 2l\u2217 )] } d\u03b8, (7) f\u03d5 = \u2212\u03bcL3R 2c2 \u222b \u03c0 0 { [(\u03c9 \u2212 2\u03d5\u0307)\u03b5 sin \u03b8 \u2212 2\u03b5\u0307 cos \u03b8 ] sin \u03b8 Gz[(1 + \u03b5 cos \u03b8)3 \u2212 12(l\u2217)2(1 + \u03b5 cos \u03b8) + 24(l\u2217)3 tanh( 1+\u03b5 cos \u03b8 2l\u2217 )] } d\u03b8. (8) 2.2 Rub-impact force Figure 1 shows the radial impact force f1 and the tangential rub force f2.f1 and f2 could be expressed as [22] f1 = (e \u2212 \u03b4)kc, (9) f2 = (f + bv)f1, if e \u2265 \u03b4. (10) Then we could get the rub-impact forces in the horizontal and vertical directions: Rx = \u2212 (e \u2212 \u03b4)kc e [ X \u2212 (f + bv)Y ] , (11) Ry = \u2212 (e \u2212 \u03b4)kc e [ (f + bv)X + Y ] . (12) 2.3 Dynamic equations Figure 2 shows a flexible rub-impact rotor supported by two couple stress fluid film journal bearings in parallel with nonlinear springs. Om is the center of rotor gravity, O1 is the geometric center of the bearing, O2 is the geometric center of the rotor, O3 is the geometric center of the journal, m is the mass of the rotor, m0 is the mass of the bearing housing, Ks is the stiffness of the shaft, K1 and K2 the stiffness of the springs which support the bearing housings, C1 and C2 the damping coefficient of the supported structure and the quadratic damping of the rotor disk, respectively, \u03c1 is the mass eccentricity of the rotor, \u03c6 the rotational angle, R the inner radius of the bearing housing and r the radius of the shaft" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002160_2011-01-1548-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002160_2011-01-1548-Figure4-1.png", + "caption": "Figure 4. Rolling element bearing kinematics and the corresponding coordinate systems.", + "texts": [ + " For each load case, the finite element equations are solved for the displacement vector of reference point represented by {di}, i = 1,2,\u20266. The support stiffness matrix can be calculated using (6) This modeling process for the mean support stiffness matrices combined with the geared rotor system dynamic model was previously validated by experimental results [17]. However, in this paper, the resultant support stiffness matrices are timevarying due to the time-varying property of rolling element bearings. The kinematics and coordinate system of a rolling element bearing is shown as Figure 4. The same nomenclature employed by Lim and Singh [1, 2, 3, 4, 5], and Liew and Lim [6] are adopted here. Raceways and rolling elements are assumed to be rigid except at the localized contact area. The Hertzian contact theory is used to calculate the normal contact forces. The bearing system is assumed to be able to rotate freely about the z-axis, and {Fbxm,Fbym,Fbzm,Mxm,Mym} is the net load vector on the rolling element bearing. The vector {\u03b4xm,\u03b4ym,\u03b4zm,\u03b2xm,\u03b2ym} is used to represent the displacements of the bearing system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000140_tmag.2007.916041-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000140_tmag.2007.916041-Figure1-1.png", + "caption": "Fig. 1. Model of a three-phase transformer with a five-limb core. Tank and windings are transparent. Clamping plates, tie bars as well as tank and yoke shielding are also shown.", + "texts": [ + " The time evolution of the magnetizing current has been qualitatively characterized in [2]. A magnetic circuit approach based on two-dimensional FEM has been used in [3]\u2013[5]. Experimental results have been presented in [6]. A method based on three-dimensional FEM to determine the wave-form of the magnetizing current in a single-phase transformer has been presented in [7]. In the present paper, this procedure is extended to the three-phase case. A three-phase, five limb core transformer is investigated. A finite element model of the transformer shown in Fig. 1 has been developed including the core, the windings, the clamping plates, the tie bars, the tank as well as the tank shielding. The model comprises of 242 550 second order hexahedral finite elements. The tank, the clamping plates and the tie bar are made of massive steel, whereas the core and the tank shielding are manufactured of laminated steel. Both ferromagnetic materials Digital Object Identifier 10.1109/TMAG.2007.916041 are nonlinear. No-load conditions are investigated with the primary windings only excited by a given sinusoidal symmetric three-phase voltage system", + " As an example, the waveforms of the three magnetizing current time-functions for the case of amperes are shown in Fig. 5. The phase shift of 120 between the three currents is obvious. The negative values are below 1.0 ampere and are therefore invisible in the plot. The distorted waveforms are due to the strong saturation illustrated in Fig. 4. The computed current waveforms have been used to carry out an eddy current analysis of the transformer with the phase-currents prescribed. The model of Fig. 1 has been simplified to concentrate on the losses in the tie bars. Only a quarter of the transformer geometry has been modeled and the tank, the shieldings and the clamping plates have been disregarded. Around each limb, one winding has been assumed carrying the magnetizing currents shown in Fig. 5. The formulation used is based on the current vector potential combined with the magnetic scalar potential [10]. Time stepping has been used and a direct iteration technique employed to treat nonlinearity [11]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001530_tmag.2010.2043827-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001530_tmag.2010.2043827-Figure3-1.png", + "caption": "Fig. 3. a) 2D electromagnetic multi-slice model with four slices. The currents within the slots are the same in all four slices. b) Structural model of the stator core stack. c) Complete mechanical model of the induction machine.", + "texts": [ + " Regarding the application to induction machines, it is therewith no longer necessary to simulate multiple electrical periods in order to get more or less acceptable results. Although theoretically only as many time steps need to be solved as there are expected frequencies in the spectrum, investigations have shown that in order to avoid numerical problems, a full electrical period of the stator current should be simulated. The coupling algorithm has been applied to estimate the mechanical excitation forces acting on the stator core stack of a skewed induction machine operating at its nominal point. Fig. 3 shows the corresponding electromagnetic and mechanical FEM models. The machine has 42 stator- and 36 rotor slots, a skewing by one stator slot pitch has been carried out on the rotor of the machine. Table I provides the basic machine data. In order to account for the rotor skewing, a 2D multi-slice model (see, e.g., [8]) with four slices has been used in the electromagnetic simulation [Fig. 3(a)]. Therewith, the magnetic flux density in the air gap has been estimated at the axial front and end of the stator core stack as well as at 1/3 and 2/3 of its axial length. In the simulation, the nominal operating point has been considered by impressing a stator current of 150 A and applying a rotational speed of 2991 rev./min to the rotor. To speed up the transient steady state computation of the eddy current problem, initial conditions for the rotor currents have been estimated by an approximate frequency domain technique as proposed in [9]", + " Comparing the numerical costs, the needed simulation time steps have been reduced altogether from 1060 electromagnetic simulation steps, consisting of 100 larger time steps for the achievement of steady state as well as 960 load steps for the three further stator current periods with a time step of 62,5 , to 260 time steps applying the present procedure and using the same 100 larger time steps for achieving the steady state as well as simulating one electrical period with a time step of 125 . Due to the elimination of the leakage effects, the accuracy of the results has been significantly enhanced, see Fig. 4. Fig. 3(c) shows the detailed 3D mechanical model of the machine, with the stator core stack in Fig. 3(b). In order to determine the electromagnetic surface force density on the air gap surface elements of the structural model, the reconstructed magnetic flux density obtained at the four slice positions in the electromagnetic model, were interpolated over the axial length of the stator core stack. Subsequently has been calculated on each surface element of the core stack applying (9) and (10). Fig. 5 shows the results for the radial component of the estimated surface force density belonging to the force spectral line at 1894" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001663_s12239-009-0049-6-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001663_s12239-009-0049-6-Figure6-1.png", + "caption": "Figure 6. Turning center of vehicle at low speeds.", + "texts": [], + "surrounding_texts": [ + "In this section, we propose an analytic model that considers returnability with an analysis of steering rack force for each directional tire force (vertical, lateral, and longitudinal). 4.1. Vertical Force A local coordinate system that follows the vehicle coordinate system is defined on the tire contact patch plane with origin A (contact patch point) in Figure 4, where D is the point where the kingpin axis crosses the contact patch plane, rc is the kingpin offset at the ground, nc is the caster trail, and \u03b4 is the tire turning angle. Moment (MV) by vertical force is defined as the cross-product of the displacement vector (rA) and the vertical axle force (FV). (1) , where As the weight of a vehicle body transfers due to lateral acceleration, the vertical axle load changes. This means that the inner axle load decreases and the outer axle load increases, even though lateral acceleration is less than 1.0 m/s2 for low-speed cornering. Lateral acceleration of a vehicle can be derived from a vehicle model with two degrees of freedom (Hwang, 2008), where the lateral acceleration in an equilibrium state is the product of yaw rate and lateral velocity, as in Equation (3). We selected a target speed of 5 km/h in order to mimic the creep speed of our target vehicle. This speed depends on engine and clutch MV = rA FV\u00d7 = nc sin \u03b4\u2013 rc cos \u03b4+( )FVi + nc cos \u03b4 rc sin \u03b4+( )FV j rA = \u2212 nc cos \u03b4 rc sin \u03b4+ nc sin \u03b4 \u2212 rc cos \u03b4\u239d \u23a0 \u239b \u239e FV = FVk characteristics of the target vehicle, and the lateral acceleration at maximum steering wheel angle was calculated to be 0.46 m/s2. Road wheel lifting effects can be ignored with little kingpin offset, since the contact patch and wheel center height do not change on steering. Therefore, since the stroke of the spring and shock absorber are very small in this system, the variation of spring and shock absorber forces do not need to be considered. Equation (2) describes the vertical forces at each road wheel, where \u03b4f represents the average value of the inner and outer steer angles, and mf g is the weight supported by the suspension spring at each front road wheel. for the inner road wheel (2) for the outer road wheel , where (3) The moment (MV) is transferred to the kingpin moment (MKV) by the projection to the kingpin axis inclination composed of the caster angle (\u03c4) and the kingpin angle (\u03c3). Equation (5) gives the real angle of the kingpin axis (\u03bb), which is the angle between the z-axis and the kingpin axis. Therefore, the restoring moment (MKV) due to vertical force is described as the product of the moment (MV) and the kingpin axis unit vector (eK), as in Equation (4). (4) , where (5) (6) Finally, Equation (7) calculates the steering rack force (Cho and Lee, 2004), since MKV is the cross product of the steering rack force (Fs) and the effective arm vector (reff), as shown in Figure 2. (7) The simulation results (Figure 5) show that at zero steering rack displacement, the total steering rack forces of the inner and outer road wheels cancel one another through left and right wheel symmetry. However, as the steering rack moves laterally, load lever arms change according to caster angle and kingpin inclination. Steering rack forces at both the inner and outer wheel increase in the restoring direction. As a result, we found that the vertical force at steering is beneficial to the restoring moment. Additionally, we considered the overturning moment obtained from the overturning coefficient (Ko), vertical force, and slip angle, but we found that it makes only a minor contribution due to low slip during low-speed turning. 4.2. Lateral Force The lateral tire force at the maximum steering wheel angle is a result of the geometric slip angle, the centripetal force from lateral acceleration, and the camber thrust from camber alteration on turning. First, the geometric slip angle is defined as follows: the normal vectors on the two front road wheels must intersect the extension of the rear axle center line (neglecting the small slip angle at the rear road wheels). The vehicle turning center (O) then becomes the geometric center of the front and rear axles for low-speed cornering (Matschinsky, 2000). In chassis system design, the inner tire turning angle (\u03b4i) is usually set to be larger than the outer tire turning angle (\u03b4o) to avoid tire scrub on cornering,which is called the Ackermann characteristic. FV = mf g H 2T --------\u2013 ay mf g+ H 2T --------ay ay = Vx 2 Wb 1 m 2Wb 2 --------- lf Cf lr Cr\u2013 Cf Cr ------------------------ Vx 2 \u2013\u239d \u23a0 \u239b \u239e -----------------------------------------------------------------\u03b4 f MKV = MV eK\u22c5 = FV nc sin \u03b4\u2013 rc cos \u03b4+( )tan \u03c4 cos \u03bb + FV nc cos \u03b4 rc sin \u03b4+( )tan \u03c3 cos \u03bb tan \u03bb = l h -- = w 2 t 2 + h -------------------- = tan 2 \u03c3 + tan 2 \u03c4 eK = w h 2 l 2 + ------------------ t h 2 l 2 + ------------------ h h 2 l 2 + ------------------ \u23a9 \u23ad \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23a8 \u23ac \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23a7 \u23ab = tan\u03c4 cos\u03bb tan\u03c3 cos\u03bb cos\u03bb\u2013 \u23a9 \u23ad \u23a8 \u23ac \u23a7 \u23ab reff Frack\u00d7 = MKV Second, vehicle lateral acceleration produces a centripetal force that acts on both road wheels. For instance, a left directional force obtained as the mass multiplied by the lateral acceleration occurs during left cornering. Third, camber thrust force (Equation 9) is included in this model due to high alteration in the lock-to-lock steering condition in a McPherson strut system. Camber angle with respect to the road varies over the range 0o ~ \u22121o at the outer road wheel and +5o ~ +7o at the inner wheel. When there is negative camber at the outer front wheel and positive camber at the inner front wheel, the lateral forces of both wheels are in the direction of the turning center. (inner wheel) (8) (outer wheel) (9) Equations (10) and (11) describe the moments around the tire contact point for geometric slip and camber angle, respectively. The tire side slip force (FS) acts on point B in Figure 4 with the lever arm (nc + nr), while the camber force (FC) acts on point A with the caster trail lever arm (nc). The kingpin axis moments in Equations (12) and (13) are obtained from the product of the moments (MS and M c ) and the kingpin axis vector (ek). The moment caused by tire side slip force is of consequence in steering returnability, because the caster trail (nc) varies considerably according to the tire turning angle. It means that the outer wheel acts on a smaller caster trail during turning, while the inner wheel acts on a larger caster trail (Figure 3). In this model, the pneumatic trail (nr) is assumed to be the same for both wheels in the low slip angle region. (10) (11) (12) (13) The simulation results (Figure 8) show that forces due to geometric slip angle are directed towards the medial plane of the vehicle for both wheels, while tire forces due to lateral acceleration are both in the same direction. The force F s , which is the sum of forces due to geometric slip angle and lateral acceleration, shows that the outer road wheel pushes the steering wheel in the returning direction, while the inner wheel pushes in the non-returning direction. The inflection of F s is found at the inner wheel occuring at around 70 mm of steering rack stroke, because the term Cf (\u03b4i \u2212 \u03b8i) is linearly decreasing. At the same time, mf ay is nonlinearly increasing due to the nonlinear steering angle, as in Figure 7. The camber thrust force caused by a large positive camber at the inner wheel rotates the steering wheel in the restoring direction, but the camber thrust force at the outer wheel is small due to the small camber alteration during turning. 4.3. Longitudinal Force In order to maintain constant speed, driving force should be applied at the wheel center position to overcome rolling resistance as well as the braking portion of the lateral force, since the longitudinal component of the lateral force applies like a braking force. Equations (14) and (15) describe the kingpin moment caused by the longitudinal force, where rC is the wheel center vector and FX is the driving force at the wheel center. (14) (15) On a wheel suspension with a fixed kingpin axis, the wheel-center offset rC remains nearly constant irrespective of tire turning angle, so that the kingpin moments of the traction force and the rolling-resistance force depend not on moment lever arm but on the longitudinal force of each road wheel. 4.4. Combined Force The total kingpin axis moment obtained by the summation FS = Cf\u2013 \u03b4i \u03b8i\u2013( ) + mf ay = Cf \u03b4o \u03b8o\u2013( ) + mf ay FC = Kf\u03b2 MS = rB FS\u00d7 MC = rA FC\u00d7 MKS = MS eK\u22c5 = FS ne nr+( )cos\u03bb MKC = MC eK\u22c5 = FCnccos\u03bb MX = rC FX\u00d7 MKX = MX eK\u22c5 = FXrwcos\u03bb of each directional kingpin moment transforms to a steering rack force with an effective steering arm vector as shown in Equation (17). The steering rack force enables the prediction of steering returnability according to its direction. If the direction of the steering rack force opposes that of a driver\u2019s steering input, then the steering wheel automatically restores to the straight-ahead position. If not, then there is no restoration. The contribution of the component force (Table 3) at maximum steering wheel angle was analyzed, and we found that steering returnability benefits from the inner wheel for vertical force, the outer wheel for side slip angle, the inner wheel for camber angle, and the inner wheel for driving force. This system, with a total steering rack force of +1,318 N in the lock-to-lock condition, can return automatically to the center position. (16) (17)" + ] + }, + { + "image_filename": "designv11_3_0002524_3.5535-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002524_3.5535-Figure3-1.png", + "caption": "Fig. 3 Precession modes.", + "texts": [ + " A projection of the path described by the nose of the vehicle on a plane perpendicular to the flight path would be as shown in Figs. 3a and 3b when the path is viewed along the direction of flight. The residual coning motion, in the same direction as the angular momentum vector (clockwise), has been called nutation by Nicolaides,7 and the retrograde precession in the opposite direction (counterclockwise) has been called, simply, precession.^ In general, the two motions exist simultaneously. For the special case of Fig. 3c, in which the vehicle is coning symmetrically about the flight path in the positive direction (direct precession), the pitch moment tends to induce a precession in the opposite direction. This condition is therefore quasi-stable, but it can persist throughout the trajectory. For the case shown in Fig. 2b, in which the coning is initially asymmetric about the velocity vector, an instability can occur in which the motion changes from direct precession to retrograde precession. This is discussed later" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000577_iemdc.2007.382825-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000577_iemdc.2007.382825-Figure6-1.png", + "caption": "Fig. 6. Artificially deteriorated bearings: (a) outer race deterioration, (b) inner race deterioration, (c) cage deterioration, (d) ball deterioration. .~~~~~~ ~ ~~ ~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~", + "texts": [ + " The bearing has eight balls (N= 8) with an approximate diameter induction motor could be identically loaded at different ofDB= 12 mm and a contact angle 0 = 0\u00b0. These bearings are speeds. made to fail by drilling holes of various radiuses with a Moreover, if the induction motor is supplied from the diamond twist bit while controlling temperature by oil network, motor current will have time and space harmonic circulation in experiments. Some of the artificially components as well as bearing fault sourced harmonics. This deteriorated bearings are shown in Fig. 6. makes it harder to determine the bearing failure effect on the stator current and therefore complicates the fault detection B. Concordia Transform Experimental Results These experiments are summarized by Fig. 7 in case of Sapigfeunyi*hse s1 H.Altedt Concordia patterns and by Fig. 8 in case of Park Patterns. obtained are used to compute stator sD-sQ and D-Q It could be seen that bearing failures cause a clear components to obtain sD-sQ (Concordia) and D-Q (Park) deformation ofthe stator current sD-sQ andD- Qtrajectories" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002838_j.proeng.2012.07.203-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002838_j.proeng.2012.07.203-Figure3-1.png", + "caption": "Fig. 3. Analytical model of three-port on/off valve", + "texts": [ + " It needs a huge time to investigate these optimal parameters by the experimental method using the real tested valve. However, in order to use the theoretical analysis, it required the model of the valve with a higher accuracy because of the critical operation of the on/off valve (PWM valve) such as repeatable operation in short period. Therefore, we proposed a precise analytical model of the on/off valve. Figure2 shows inner construction of the three-port type on/off valve that have been used in this study. The valve consists of a solenoid, an armature with a spring. Figure3 shows the analytical model of the three-port type on/off valve. The model of the electric circuit of the solenoid is used as an electric coil and resistance. The armature has a mechanical system to control the fluid flow so that the supply ports and exhaust ports can\u2019t be opened at the same time. There is an overlap that the valve cannot open and close at the point of a certain displacement (= xmin) of the armature. The displacement of the armature also has the maximum limit (= xmax) of a moving area" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001225_tase.2009.2020994-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001225_tase.2009.2020994-Figure1-1.png", + "caption": "Fig. 1. Tangent plane at the reference point of sampling.", + "texts": [ + " The tactile data are specially arranged so we can estimate the principal curvatures and locate the principal directions. We estimate the two principal curvatures and at the reference point from normal curvatures in different tangent directions. The angle between one of these tangents and one principal direction also determines the other principal direction due to their orthogonality in the tangent plane. So there are essentially three unknowns: . This suggests that we need to measure at least three normal curvatures, say, , and in the tangent directions , and (shown in Fig. 1), respectively. How to measure these curvatures will be described in Section III-C. The two principal directions are and . The angle is from to , while the angles and are from to and , respectively. Here and are easily determined from the tangents. The normal curvatures can be expressed in terms of the two principal curvatures [41, p. 137] (2) From (2), we obtain (3) (4) In the special case that , two possibilities arise. a) . So, the reference point is umbilic with constant normal curvature. Every direction in the tangent plane is a principal direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002420_1.4001827-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002420_1.4001827-Figure3-1.png", + "caption": "Fig. 3 Measured and predicted temperatures rise \u201eTDE\u2212Tamb\u2026 at the OD of drive end bearing versus rotor speed. Operation without and with a cooling stream at \u00c850 l /min. Ambient and cooling stream temperatures, Tamb=TCo=21\u00b0C. Heater off \u201ecold", + "texts": [ + " Presently, thermal expansion of the whole test rig casing is not considered. In the model, an empirically derived thermal mixing coefficient =0.65 represents best the gas flow and thermal energy balance at the conjunction of the top foil leading and trailing edges 2 . 2.1 GFB Operation With Cold Rotor (No Heating). In test condition 1 in Ref. 1 , the rotor speed varies from 0\u201330 krpm, the cartridge heater is switched off inactive , and thus the rotor OD surface is cold at ambient temperature. Figure 3 shows the predicted and the measured temperature rise TDE\u2212Tamb at the outboard of the drive end bearing versus rotor speed for operation without and with a forced cooling stream at 50 l/min and temperature of 21\u00b0C, equal to ambient temperature, TCo=Tamb. A fraction of the rotor weight 6.5 N acts on the bearing. The inset cartoon shows the GFB noting the orientation of the top foil spot-weld with respect to the vertical gravity plane. The test data represent the arithmetic mean from four temperature measurements around the bearing circumference; maximum and minimum values are also depicted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001707_bsn.2010.46-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001707_bsn.2010.46-Figure2-1.png", + "caption": "Fig. 2. Top view of club face and ball at impact showing face angle.", + "texts": [ + " While examining the figure note the coordinate system that will be used throughout this paper. Positive z points upward perpendicular to the ground, positive x denotes the intended ball path and positive y points to the golfer while remaining perpendicular to x and z. There are many parameters of a golf swing that affect the trajectory of the golf ball. The goal of our model is to identify parameters that address precision and repeatability of the swing. We focus our attention on the following most critical parameters: \u2022 Face angle (\u03c8) at impact (Figure 2) \u2022 Loft angle (\u03b8) at impact (Figure 3) \u2022 Lie angle (\u03c6) at impact (Figure 4) \u2022 Velocity throughout swing \u2022 Location (x\u2032, y\u2032) of impact on club face (Figure 4) \u2022 Motion path immediately surrounding impact1 \u2022 Tempo: Proportion of back-swing duration to forwardswing duration By examining each of Figures 2, 3 and 4 it can be observed that the most predictable path of travel for the golf ball will occur when loft, face and lie angles are all zero. In fact the motion is so sensitive to error that a \u03c8 = \u00b13 degree face angle will result in an error greater than 15cm for a 3m putt" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001964_j.optlaseng.2010.09.017-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001964_j.optlaseng.2010.09.017-Figure1-1.png", + "caption": "Fig. 1. Laser heating situation and coordinate system.", + "texts": [ + " The heat transfer equation in relation to the laser heating process can be written as [16] rDE Dt \u00bc \u00f0r\u00f0krT\u00de\u00de\u00feSo \u00f01\u00de where E is the energy gain by the substrate material, k the thermal conductivity, and So the volumetric heat source term and it is So \u00bc Iod\u00f01 rf \u00deexp\u00f0 dz\u00deexp r a \u00feb 2 f \u00f0t\u00de Io is laser peak intensity, d the absorption depth, a the Gaussian parameter, rf the surface reflectivity, b the density parameter, and x and y are the axes when the laser beam scans the surface along the x-axis. The laser beam axis is the z-axis (Fig. 1). Fig. 2 shows the normalized laser power intensity distribution along the x-axis. In the case of a moving heat source along the x-axis with a constant velocity U, energy gain by the substrate material satisfies rDE Dt \u00bc r @E @t rU @E @x \u00f02\u00de or rDE Dt \u00bc r @\u00f0CpT\u00de @t rU @\u00f0CpT\u00de @x \u00f03\u00de Combining Eqs. (1) and (3) yields r @\u00f0CpT\u00de @t \u00bc \u00f0r\u00f0krT\u00de\u00de\u00ferU @\u00f0CpT\u00de @x \u00feSo \u00f04\u00de Since the laser scanning speed remains constant, the heating situation can be considered as a steady, in which case the term r(q(CpT)/qt) in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002736_tmag.2012.2196502-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002736_tmag.2012.2196502-Figure6-1.png", + "caption": "Fig. 6. Analyzed model.", + "texts": [ + " Also increased in the right and left yokes under tensile stress. The distributions were similar to the distributions because the flux condition was almost alternating in the ring-core model. Since the tendency of the magnetic characteristic distribution depending on the residual stress agrees with the measured one (Fig. 2), it can be said that the SCES modeling is applicable to magnetic characteristic analysis considering stress effect in core materials. The developed method was applied to analysis a PM motor model core as shown in Fig. 6. Table II shows the conditions used in the analysis. The motor model core has 4 poles and 36 slots. Fig. 7 shows the distributions of the residual stress in R.D. and T.D. The residual stress in the stator core was measured by using the X-ray stress measurement system [4]. The rolling direction agrees with the X-direction. However the local easy axis depends on the residual stress distribution. The compressive stress in R.D. and T.D. was distributed mainly in the back yoke and the largest tensile and compressive stress occurred around the caulking" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003565_1350650112468071-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003565_1350650112468071-Figure4-1.png", + "caption": "Figure 4. Sketch of the rectangular plate element.", + "texts": [ + " The bending and shear deformation of the top foil are coupled in the thick plate model. The implementation of foil structure is shown in Figure 3. The bump foil is considered as linear spring and the bump structual stiffness per unit area is estimated according to the geometrical values of bump foil. For the Kirchhoff plate theory, the governing equation for derermining the transverse w (in the z direction) is D @4w0 @x04 \u00fe 2D @4w0 @x02@y02 \u00feD @4w0 @y04 \u00bc q0 \u00f08\u00de The rectangular elastic plate element is shown in Figure 4. The stiffness matrix \u00bdke K of top foil element is derived by using virtual work and is shown in the Appendix 2. As in the case of thick plate, the shear deformation and rotary inertial effects must be included. The relationships between the transverse shear strains xz, yz and the displacements are xz \u00bc @u @z \u00fe @w @x , yz \u00bc @v @z \u00fe @w @y \u00f09\u00de The energy expressions for a thick plate are Ue \u00bc 1 2 Z A h3 12 f gT\u00bdD f gdA\u00fe 1 2 Z A hf gT\u00bdDs f gdA \u00f010\u00de at SIMON FRASER LIBRARY on June 15, 2015pij.sagepub" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002017_pedstc.2010.5471830-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002017_pedstc.2010.5471830-Figure2-1.png", + "caption": "Fig. 2. a) Stator winding flux axes b) Rotor winding flux axes", + "texts": [ + " In section II, a method for modeling a faulty three phase induction motor (when it loses one of its feeding phases) is presented. By using this model, a new method of vector control is presented in section III. The Performance of the presented method is checked by computer simulation in section IV. Suppose that a phase cut out fault is occurred in the phase \u201cc\u201d of a three-phase drive system, as shown in fig 1. Assuming sinusoidal waveform for the spatial distribution of the windings, Stator and rotor winding flux, axes can be shown as fig 2. If the angle between ds-axis and as-axis is \u201c\u03b8o\u201d, d and q components of flux can be written as follows: [ ] \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 += bs as oods \u03d5 \u03d5 \u03c0\u03b8\u03b8\u03d5 )3/2cos()cos( (1) [ ] \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 += bs as ooqs \u03d5 \u03d5 \u03c0\u03b8\u03b8\u03d5 )3/2sin()sin( (2) Modeling and Vector Control of Unbalanced Induction Motors (Faulty Three Phase or Single Phase Induction Motors) M. Jannati, and E. Fallah T 1st Power Electronic & Drive Systems & Technologies Conference 978-1-4244-5971-1/10/$26.00 \u00a92010 IEEE 208 The transformation vectors \u201cd\u201d and \u201cq\u201d can be defined as follows: [ ])3/2cos()cos( \u03c0\u03b8\u03b8 += ood (3) [ ])3/2sin()sin( \u03c0\u03b8\u03b8 += ooq (4) The transformation vectors must be perpendicular, so we have: 6/ 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003105_j.ijsolstr.2011.09.017-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003105_j.ijsolstr.2011.09.017-Figure4-1.png", + "caption": "Fig. 4. Generalized constrained Ziegler column problem.", + "texts": [ + " (50), to investigate which constraints are the most critical for the Ziegler column, that is, which constraints ensure the destabilization of the Ziegler column for a given a value at the lowest loading force ps 1. Generalizing the so-called isochoric condition (Eq. (50)) considered above, the following constraint: a1h1 \u00fe a2h2 \u00bc 0 \u00f053\u00de can be introduced. Hereafter, parameter a1 is assumed not to be nil, so that Eq. (44) reads: h1 \u00fe ah2 \u00bc 0 \u00f054\u00de with a \u00bc a2 a1 . The technological device applying the constraint given in Eq. (54) is described in Fig. 4. Ignoring the weight of the different parts of the structure, the bar BC is supported by a beam (D), with a length ideally infinite. The lateral deviation of a given point D of the beam along the axis (Ay) is prevented. Then the following holds: h1 \u00fe BD L h2 \u00bc 0 \u00f055\u00de where the algebraic quantity a \u00bc BD L can take positive or negative values, its absolute value being less or greater than 1, according to the relative position of point D with respect to the segment [BC]. For the p values within the range ps 1; p s 2 , the set I gathering all vectors x ensuring that the quantity tx Ksx is negative is not reduced to the nil vector" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001514_tmag.2009.2018676-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001514_tmag.2009.2018676-Figure1-1.png", + "caption": "Fig. 1. Permanent magnet spherical wheel motor.", + "texts": [ + " Index Terms\u2014Current compensation, curve fitting method, permanent magnets electric machines, spherical wheel motor. I. INTRODUCTION T HE spherical wheel motor is remarkable electric machine, because the motor is very useful for multi-degrees of freedom system, for instance manipulator, humanoid [1]. A rotor of the spherical wheel motor can tilt and rotate as using alignment torque between the magnetic field produced by magnet and magnetomotive force (MMF) created by coils on 3-D space. Therefore, the spherical wheel motor is called the three degrees of freedom actuator [4]. Fig. 1 is spherical wheel motor researched in this paper. The spherical wheel motor consists of 4 permanent magnets on rotor and 12 coils mounted on inner-surface of stator. Each coil is made of 400 turns and is wound around a bobbin. For positioning a shaft, each coil is excited with current function as using alpha degree, beta degree, and theta degree which are defined desired rotor position. A rotor of the proto-type spherical wheel motor can tilt up to 18 degrees and rotate at 160 r/min. Shaft position is detected three rotary encorders mounted on guide frame" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000654_s11071-009-9482-3-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000654_s11071-009-9482-3-Figure7-1.png", + "caption": "Fig. 7 Collision model within the bearing clearance", + "texts": [ + " In addition, the lateral static force f0 is added to simulate the eccentricity of the bearing in the radial clearance caused by an assembling error. The terms of ip\u03c9y\u03071,\u2212ip\u03c9x\u03071 are also added to consider the gyroscopic moment of the disk. The notations ks1, ks2 are the spring coefficients of the rotor, and can be written as ks1 = 3EsIsl a2b2 , ks2 = 3EsIs ab2 = l a ks1. (2) Here, Es is Young\u2019s modulus of the rotor, Is is the moment of inertia of area. 3.2 Modeling of collision in bearing The collision between the bearing and the casing is explained in Fig. 7. This treatment follows the method introduced by Li and Paidoussis [6]. In Fig. 7 where the bearing is sketched small relatively to the casing for the explanation, r2 is the radial displacement of the bearing, vn and vt are the velocity components of the bearing in the normal and the tangential direction, respectively. In numerical calculation, the bearing orbit is given by a succession of discrete points. Suppose that the bearing moves towards the casing wall and P1 is the discrete point just one step before the collision. As the next discrete point P2\u2212 passes a little through the wall, the position is transferred to P2+ as following condition under the assumption that it bound at P12", + " This difference is explained as follows. In the theoretical model, it is assumed that the bearing always bounce off when it collides the casing. On the other hand, in the experiments, the bearing probably slides on the inner wall without bouncing when the harmonic response occurs. As an example, the subharmonic resonance of order 1/2 in Figs. 5 and 9 is analyzed theoretically in this section. In order to derive its solution, the Harmonic Balance Method (HBM) [8] is used in this study. As the collision model shown in Fig. 7 is difficult to analyze theoretically, an equivalent spring and damper model is introduced to simulate the collision in the clearance. 4.1 Equivalent spring and damper model of collision Replacing the collision of the bearing with the casing to the instantaneous viscoelastic deformation, the radial contact force Fn at the collision is represented by Fn = \u23a7\u23aa\u23a8 \u23aa\u23a9 kbn(r2 \u2212 \u03b4) + cbn(x\u03072 cos \u03b8 + y\u03072 sin \u03b8) (r2 \u2265 \u03b4), 0 (r2 < \u03b4), (7) where kbn, cbn are the spring and the damper coefficients, \u03b8 = tan\u22121(y2/x2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003830_icmech.2013.6518544-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003830_icmech.2013.6518544-Figure3-1.png", + "caption": "Fig. 3: body frame offset", + "texts": [ + " The details of contact modeling and simulation algorithm are in [17]. The modeled reaction forces suffer from peaks, so that Kalman filter is used for smoothing the modeled reaction forces. The coordinate frames are shown in Fig. 2. All the measurements and calculation are in the world frame. The transformation is done using the rotational matrix obtained by the author in [18]. The body frame has an offset offsetx , The role of this offset parameter is to place the center of the support polygon exactly below the center of mass of the robot as shown in Fig. 3. According to the figure it has a negative value. Kalman filter parameters are listed in Table II. , ,Q R and the low pass filters constants are chosen by trial and error. The estimated ZMP, CoM states and the disturbance for the x \u2212 direction is depicted in Fig. 4. The estimated ZMP (Fig. 4.a) shows its reliability and accuracy compared to the reference ZMP and the ZMP based on modeled and filtered force. The estimated CoM position trajectories and reference desired ones are in the body frame as expressed in the world frame, the simulated trajectories are in the CoM frame as expressed in the world frame" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002594_j.ces.2011.07.026-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002594_j.ces.2011.07.026-Figure1-1.png", + "caption": "Fig. 1. Scheme for the principal parts of the CLC-7000.", + "texts": [ + " Such suspensions are known to exhibit non-Newtonian shear-thinning or shear-thickening behavior, depending on the solids content and flow conditions (Roper and Attal, 1993). It is known that the rheology of these suspensions, also called coating colors, is a key parameter for controlling the runnability of coating applications, especially at high coating speeds (Alonso et al., 2003). Because runnability problems were not encountered in this work, a more detailed rheological characterization of the suspension used, with more flexible rheometers, was not performed. A Cylindrical Laboratory Coater model 7000 (CLC, see Fig. 1) was used in order to apply the suspension on paper. The CLC allows for reproducible coatings using commercial parameters at high speed (maximum at 2200 m/min): suspension viscosities and drying temperatures. Although the blade loading is different at the pilot and industrial scale, the shear forces are quite similar and the coating weights can be adjusted by changing the pressure exerted on the blade on the paper (Egorova, 2006; Horsfield, 2003). Wood-free paper strips of 0.75 3 m2 at an average weight of 53 g/m2 were used as paper base for the applications; to avoid any interactions between lignin and laccase (Leonowicz et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003536_s1064230712020062-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003536_s1064230712020062-Figure1-1.png", + "caption": "Fig. 1. Two level architecture of the self organizing flood forecasting system (STAFF).", + "texts": [ + " It is worth noting that the research and development in self organizing MAS technology are still in the initial stage, and much effort remains to be made in order to obtain a mature technology and tools for its support. A characteristic example of a self organizing MAS is the system for forecasting floods in the Rhone basin (STAFF) [1, 2]. It includes several stations installed over the river basin that forecast local variation in the water level. DOI: 10.1134/S1064230712020062 392 JOURNAL OF COMPUTER AND SYSTEMS SCIENCES INTERNATIONAL Vol. 51 No. 3 2012 The software installed on each station has the two level multiagent architecture depicted in Fig. 1. The lower level includes sensors that register hourly variations in the water level, and these sensors can provide erroneous data. A station can use data obtained from thousands of sensors. A sensor agent takes the vari ation in the water level measured by the sensor i. This agent is responsible for calculating the weight with which the quantity for the time will be taken into account in the forecast function calculated by the upper level agent Aj ( ) at the current time to forecast the variation in the water level for the next hour, that is, for tk: (1", + " A conflict triggers the self organization process on the set of lower level agents of the MAS in which A1 plays the role of the upper level agent (this process is described below). Upon obtaining the new values of the weights, the agent A1 calculates the next water level forecast for the time . 2. The upper level agent A2, which makes its forecast for the same time , compares its forecast with the new forecast made by A1 for . If the difference exceeds a certain threshold, then A2 starts the self organization process on the set of its lower level agents. Next, these procedures are carried out in a loop by the agents A3 and A4 within their MASs installed on the station. Figure 1 illustrates this cooperation process between the upper level agents by arrows. As a result, all agents acting at the metalevel can be involved in self organization. Now, consider the idea of the self organization process that is carried out by each station MAS on the set of its lower level agents. Let Si be the measurement of the sensor i of this station, \u03c9i be the current weight with which this measurement is fed to the input of the upper level agent. This weight cannot be negative; if \u03c9i becomes negative in the course of computations, it is set to zero (see (1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001895_978-3-642-32448-2_11-Figure11.8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001895_978-3-642-32448-2_11-Figure11.8-1.png", + "caption": "Fig. 11.8 Modal analysis of the robot", + "texts": [ + " In order to identify the mechanical model parameters of the robot several experiments are performed including \u2022 modal analysis, \u2022 stiffness measurement within the working volume of the robot, \u2022 stiffness measurement of the structural robot components. In the following sub-sections, the investigations are presented and the consequential model adaptations are implemented. The modal analysis is used to determine eigenfrequencies, eigenforms and modal damping of the robot structure. The robot structure is excited by a hammer impulse at a defined position. The response is recorded with a tri-axial acceleration sensor, which the transfer functions are derived from for all 109 measuring points. Figure 11.8 shows the robot structure and the measuring grid. 11 Analysis of Industrial Robot Structure and Milling Process Interaction 255 Based on the transfer functions, the eigenforms, eigenfrequencies and modal damping is extracted. The robot KUKA KR 210 [17] has six dominant eigenfrequencies below 100 Hz. In Table 11.1, these frequencies are summarized and explained briefly. Measuring of the static stiffness in the working space is performed in x, y, and zdirection at nine positions within the relevant working area of the robot" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002006_j.jmatprotec.2012.02.016-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002006_j.jmatprotec.2012.02.016-Figure1-1.png", + "caption": "Fig. 1. Schematic of gas metal arc welding experimental system.", + "texts": [ + " Schematic diagram showing measurement of characteristics of a In this paper, the angle between the electrode wire and the workpiece is acute and we focus on the impact on workpiece stage of globular metal transfer at various welding velocities in the GMAW process. The impact location, diameter, kinetic energy and rebound percentage of droplets are measured by the threedimensional information which is obtained by high speed video photography with a laser source. Furthermore, the influence of the droplet behavior on the weld morphology is also analysed. 2. Experimental The schematic of gas metal arc welding experimental system is shown in Fig. 1. The system was composed of three parts, including a welding system, a frame acquisition system and a welding current and voltage collection system. The welding system: a welding machine (YD-500AG) with 500 A current rating was used in droplet. (a) Flight distance and (b) impact location and diameter. J. Feng et al. / Journal of Materials Processing Technology 212 (2012) 2163\u2013 2172 2165 2166 J. Feng et al. / Journal of Materials Processing F v r 1 I t r a L b u f T w r = D 2 (4) ig. 7. Welding current and voltage at different welding velocities", + " he computer was used to measure the characteristics of a droplet ith image processing software (Adobe Photoshop) and to collect Technology 212 (2012) 2163\u2013 2172 welding current and voltage with a general-purpose, procedural programming language (LabVIEW). Bead-on-plate welding was carried out on the base metal (E36 steel) in the dimension of 150 mm \u00d7 60 mm \u00d7 16 mm. The electrode wire used was ER50-6 of 1.2 mm in diameter. The composition of the base metal and electrode wire are listed in Tables 1 and 2, respectively. The Angle (\u02db) between the GMAW torch and the workpiece and the angle (\u02c7) between the high speed video and the workpiece are shown in Fig. 1. All the processing parameters are shown in Table 3. The angle (\u02c7) was set as 10\u25e6 and 30\u25e6 to acquire more information of the welding process. The surface of the sample was cleaned by acetone before welding. The influence of welding velocity on the impact behavior of globular metal transfer was studied by high speed video photography with a laser source during the GMAW process. The various welding velocities were 0.4 m/min, 0.8 m/min, 1.2 m/min, 1.6 m/min and 2.0 m/min. The impact behavior of a droplet was acquired by analyzing consecutive frames of the video photograph during the welding process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000947_tmag.2009.2012590-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000947_tmag.2009.2012590-Figure3-1.png", + "caption": "Fig. 3. Finite-element analysis results for the mutual flux linkage of the inactive phase lagging 240 (electrical) of the prototype 6/4 SRM as a function of current and angle from aligned position for single-phase excitation.", + "texts": [], + "surrounding_texts": [ + "Finite-Element-Based Estimator for High-Performance Switched Reluctance Machine Drives\nIakovos St. Manolas, Antonios G. Kladas, and Stefanos N. Manias, Fellow, IEEE\nFaculty of Electrical and Computer Engineering, National Technical University of Athens, Athens15121, Greece\nThe majority of existing models for estimation and control of switched reluctance machines (SRMs) ignore the dynamic effects of mutual coupling between different phases. Although this may be acceptable for some common industrial applications, such a simplification may degrade significantly the performance of the drive and the precision of the estimated quantities. In this work an enhanced representation of the SRM is introduced, by defining an appropriate finite element model of the machine. This model enables accurate estimation of rotor position and electromagnetic torque, for any current configuration. The necessity as well as the feasibility of this approach is illustrated in the case of a three-phase, 6/4 SRM. Torque and position estimation accuracy provided by the proposed methodology is validated by measurements on a prototype three-phase 6/4 SRM.\nIndex Terms\u2014Angular position estimation, electromagnetic torque estimation, finite-element-based estimator, switched reluctance machine (SRM) drives.\nI. INTRODUCTION\nT HE recent advances in power electronics and control theory have encouraged switched reluctance machine (SRM) drives as an interesting alternative in variable speed applications. SRMs are doubly salient machines, in which torque is produced when the rotor moves to the position maximizing the inductance of the excited winding. The per phase voltage equations can be formulated as\n(1)\nwhere , , , , and stand for phase applied voltage, resistance, current, flux linkage and rotor position, respectively. Due to the inherent iron saturation, the self and mutual flux linkages are nonlinear functions dependent on rotor position and phase currents.\nThe procedure of accurate modelling and sensorless estimation under all operating conditions in such a nonlinear system is very demanding [1]\u2013[7]. The common approach is the formulation of a model for the SRM based on flux linkage data. There are two ways for obtaining these flux linkage data: either by experimental measurements [1] or by field analysis of the machine in question. The SRM model may be either detailed, including the full dynamic effects of self and mutual flux linkages, or reduced, where only the self flux linkages are considered. In the majority of cases, the SRM is modeled with data produced via unipolar current single phase excitation for various rotor positions. However, it is very usual that two adjacent phases are conducting simultaneously, especially in high-speed applications, where the current tails are significant. In these cases, the model of the SRM fails [1], [2] and, as a result, significant errors arise both in rotor position and torque estimation.\nIn this paper, the SRM in question undergoes a detailed parametric finite-element analysis [8] for various rotor positions, in which only one phase is excited with various current levels.\nManuscript received October 07, 2008. Current version published February 19, 2009. Corresponding author: I. St. Manolas (e-mail: imanolas@central.ntua. gr).\nColor versions of one or more of the figures in this paper are available online at http://ieeexplore.ieee.org.\nDigital Object Identifier 10.1109/TMAG.2009.2012590\nThe magnetostatic problem is solved for 46 rotor positions and for nine current levels for each position, i.e., 414 simulations are carried out. The resultant self and mutual flux linkages and the produced electromagnetic torque are classified in a look-up table, which constitutes the dynamic model of the SRM for single-phase excitation. The above results are presented graphically in Figs. 1\u20134. The goal of this work is to propose a way of acquiring a satisfactory estimation of the rotor angular position and the produced electromagnetic torque, which will also address the case of two phase excitation of the SRM based only on the existing flux linkage data.\nIn general, the estimation of the angular rotor position is based on the flux linkage data, which constitute the SRM\u2019s dynamic model. It should be noted though, as already mentioned, that these data produce accurate position estimation only for the case of single phase excitation. If more than one of the SRM\u2019s phases are conducting simultaneously, as favored in high speed applications, the preceding model fails and new flux linkage data need to be collected. This is a very complicated\n0018-9464/$25.00 \u00a9 2009 IEEE", + "and exacting procedure that demands increased computational and memory resources.\nIn this work, an innovative algorithm is proposed, which allows for a satisfactory estimation of the angular rotor position both for single- and multi-phase excitation, based only on the existing self and mutual flux linkage data. To achieve this goal, an appropriate projection of the observed flux linkages to the existing data set needs to be introduced.\nThe first step of the algorithm is the identification of the primary and the secondary active phase. The phase current is not a suitable criterion, since the variable reluctance may in some cases change the roles of the phases. As primary active phase, we choose the phase with the maximum observed flux linkage, whereas as secondary active phase the one with the intermediate flux linkage level.\nBy examining the estimation results produced by the commonly used method of linear interpolation based only on the self-flux linkage data, one can perceive that the estimation errors that arise are significant, especially around the aligned position for the primary active phase. These errors are mainly imputed to two factors: the magnetic saturation and the interaction between the two active phases.\nIn Fig. 5, the resultant field distribution of a 6/4 SRM for two phase excitation is depicted. The resultant flux linkages from this field distribution can be interpreted as a suitable superposition of the existing data of the self- and mutual flux linkages for single phase excitation, as shown in Fig. 6. Thus, the flux linkage of the phase that is positioned at zero degrees in Fig. 6 includes two contributions: one that can be attributed to the self linkage of the primary active phase and one that can be attributed to the mutual linkage of the secondary active phase. Since the system is highly nonlinear, a simple superposition of the existing quantities for single-phase excitation is not appropriate.\nTherefore, a different approach needs to be adopted. The goal is to project the observed flux linkage of the primary active phase to the known flux linkage data set. For this purpose, the mutual flux linkage of the secondary active phase is used. The projected value of the self flux linkage can be evaluated using the following equation:\n(2)", + "where , , stand for the projected self flux linkage of phase A, the observed flux linkage and the mutual flux linkage of phase A generated by phase B, respectively. The is the weight function that will ensure an appropriate projection and will cover the cases of single and multi-phase excitation. This weight function is given by the formula\n(3)\nwhere and is the observed current of the primary and the secondary active phase, respectively. Finally, the projected self flux linkage value is interpolated with the existing flux linkage data to acquire an estimation for the angular position of the rotor.\nThe above algorithm is used to address the estimation problem for all cases, even those in which the commonly used method fails. The SRM undergoes a finite element analysis for various rotor positions and current configurations to define the observed flux linkages. In most cases a 2-D analysis provides sufficient accuracy, while, in cases of unaligned rotor position and those involving highly saturated iron, a 3-D analysis is necessary. The estimated rotor positions are shown in Table I.\nGiven the above algorithm, which provides a satisfactory estimation for the rotor angular position, the next step is to acquire an estimation for the produced electromagnetic torque. Again in this case, only the torque data for single phase excitation is required.\nThe estimation of the produced electromagnetic torque for each conducting phase is obtained by using the rotor angular position acquired with the rotor position estimator analyzed above. The total produced electromagnetic torque is the simple superposition of the produced torque of all phases. This procedure is visualized in Fig. 8. In Table II, the values of the produced torque for all the cases of Table I evaluated by accurate finite-element analysis are compared to those estimated by the proposed method.\nThe accuracy of the torque and position estimation provided by the proposed methodology is validated by measurements on the prototype three-phase, 6/4 SRM depicted in Fig. 7. The most significant estimation errors arise near the aligned and the unaligned rotor positions, which are illustrated in Fig. 9. This figure compares the measured rotor position to the estimated" + ] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure12-1.png", + "caption": "Fig. 12 Balanced crank-slider mechanism synthesized from the CRCM configuration with two CRCMs near the base", + "texts": [ + " In these 1DOF crank-slider mechanisms, 2 depends on 1. This relation is easy to find from the second equation of r2 and its derivative: r2,y = l1 sin 1 + l2 sin 2 = h 48 Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use W w A a S r T p s t c p m s t t l t c s s t C a F t p J Downloaded Fr r\u03072,y = l1\u03071 cos 1 + l2\u03072 cos 2 = 0 49 ith \u03072= \u03071+ \u03072, \u03072 can then be written as \u03072 = \u2212 l1 cos 1 l2 cos 2 \u2212 1 \u03071 50 ith 2 = sin\u22121 h \u2212 l1 sin 1 l2 51 s an example, the configuration of Fig. 12 is taken for which the ngular momentum writes from Eq. 37 : hO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a \u03071 + I2 + m2l2 2 + m2 l2 2 + k2I1,b \u03072 52 ubstituting the kinematic relations of Eqs. 50 and 51 then esults in hO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a + I2 + m2l2 2 + m2 l2 2 + k2I1,b \u2212 l1 cos 1 l2 cos 2 \u2212 1 \u03071 53 he single equation of the reduced inertia is now dependent on the osition of the mechanism. This equation also holds for an uncontrained balanced double pendulum moving along the same trajecory, although then there are two input angles where each has a onstant reduced inertia" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000370_bf01228610-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000370_bf01228610-Figure6-1.png", + "caption": "Fig. 6. The loci of F1 = 0 and F2 = 0 in the v, r/plane for a = + 1. These loci, determined from Equations (23) and (24) respectively, relate to theextremal values of C given by OC/OJ7 ~ C , = 0 and OC/Ov ~- C,=0. In particular, the locus for F2 = 0 is the subset of the F1 = 0 locus composed of the diagonal straight", + "texts": [ + " - ac,(1 + c 2) G = Q - = o ( 2 3 ) s~Gs, s,G and. for C~=O. 2c~ a04C, - ac.(1 + 4 ) = 0 2c~ s~s~c. - at.(1 + e2~) = o (24) 432 D O N A L D L. HITZL A N D MICHEL HI~NON 0,/ k.) 6 O, mb 6 d ~ 6 0 ! It ! ! 0 ~5 -.-- I, 0 I I Q I I ! b ! ,2o t~ r ~ cD r Ox r~ to O a~ ,.2o om,,~ ,,.{3 \"~ , u d ,m ~ d o c ~ r~ to O kO q_) q ..r ..r oq GENERATING ORBITS FOR SECOND SPECIES PERIODIC SOLUTIONS ! I! I I I P ~ ! 433 ~o raO r ~ r~ ~0 O~ O ~ o ~ nl ro -6 ~ ~,,,q tD r162 ,.o g~ O r,r O ~5 434 DONALD L. HITZL AND MICHEL HI'NON These loci are shown in Figure 6 for both a = + 1. Again, the invariance properties of lines. 4. Criticality Condition Consecutive collision orbits located at extremal values of C can now be obtained analytically as follows. The orbit must be located at an extremal value of C together with the constraint that the orbit also satisfies the timing condition Fo=0 of (11). Thus we set e~ [c(~' v) + ,~Fo(~, V)] = 0 [c(~, ~) + ZFo(~, ~)1 = 0 (25) yielding an equation = - C . F o = O G A C, Fo, (26) after the unknown Lagrange multiplier 2 is eliminated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002962_ja206136h-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002962_ja206136h-Figure1-1.png", + "caption": "Figure 1. Diagram of SECM inside a drybox.", + "texts": [ + " All SECM and other electrochemical measurements were carried out with a CHI 920C SECM station bipotentiostat (CH Instruments, Austin, TX). The SECM scanning head and cell were placed inside the drybox and connected to the controller outside the glovebox via a feed-through. This was necessary to avoid contamination of the solution with water and oxygen. A partially inflated inner tube was placed beneath the SECM to minimize vibrations from the glovebox blower, and a metal plate was placed on top of the tire as a stand support. A diagram of the setup is shown is Figure 1. Electrodes. Platinum (99.99%) 10 \u03bcm diameter wire from Goodfellow (Devon, PA) was used to fabricate the SECM electrodes by procedures described elsewhere.13,14 The glass surrounding the tip was made as small as possible by careful polishing, holding the tip at an angle with respect to the polishing disk. Making such a tip and careful alignment with respect to the substrate were necessary to obtain the very small d/a values required in the measurement. Details about the tip construction and simulations of the tip behavior with geometric variations will be published elsewhere" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure11-1.png", + "caption": "Fig. 11. (a) Double chain with five pairs: three non-intersecting pairs followed by two intersecting pairs. (b) The edge vectors and the inclination angles.", + "texts": [ + " (4) because it is again automatically satisfied. Using the same approach as in Section 1, we can find that the mobility conditions are p1 \u00fe p2 \u00fe p3 \u00fe p4 \u00fe p5 \u00bc 0 \u00f027\u00de and q1 \u00fe q2 \u00fe q3 \u00fe q4 \u00fe q5 \u00bc 0 \u00f028\u00de as indicated in Fig. 9b. Two snap shots of the motion sequence of a physical model based on this particular type are shown in Fig. 10, which confirms our findings. Other possible combinations for double chain assembly with five pairs exist. For example, it can have two intersecting pairs followed by three non-intersecting pairs, as shown in Fig. 11. Let us now investigate the mobility of this assembly. To preserve the parallelogram condition, the diagrams shown in Fig. 11b can be drawn. There must be a1 \u00fe a2 \u00fe a3 \u00fe a4 \u00fe b5 \u00fe h \u00bc 2p; \u00f029\u00de and b1 \u00fe b2 \u00fe b3 \u00fe b4 \u00fe a5 h \u00bc 2p: \u00f030\u00de Using the same argument for the immobile double chain assembly consisting of five non-intersecting pairs, it can be concluded that this assembly is also immobile. A physical model, shown in Fig. 12, confirms that this is true. For all of the double chain assemblies made from five pairs, we can determine their mobility using the same approach. The result is given in Table 1. Note that for an immobile double chain the total number of non-intersecting pairs is always odd" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001144_j.snb.2008.12.041-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001144_j.snb.2008.12.041-Figure1-1.png", + "caption": "Fig. 1. Schematic drawing of the experimental setup. A sessile droplet is positioned between two parallel electrodes with various types of insulation as described in the text. The contact angle is measured by means of video microscopy. This figure is not to scale.", + "texts": [], + "surrounding_texts": [ + "R B U a b\na\nA A\nK E D C H S\n1\nd e o d g o l\nc ( v\nc\nw v s\n0 d\nSensors and Actuators B 144 (2010) 380\u2013386\nContents lists available at ScienceDirect\nSensors and Actuators B: Chemical\njourna l homepage: www.e lsev ier .com/ locate /snb\neversible electrowetting on silanized silicon nitride\nrian P. Cahill a,\u2217, Athanasios T. Giannitsis a, Raul Land a,b, Gunter Gastrock a, we Pliquett a, Dieter Frense a, Mart Min a,b, Dieter Beckmann a\nInstitute for Bioprocessing and Analytical Measurement Techniques, Rosenhof, 37308 Heilbad Heiligenstadt, Germany Tallinn University of Technology, Ehitajate tee 5, 19086 Tallinn, Estonia\nr t i c l e i n f o\nrticle history: vailable online 31 December 2008\neywords: lectrowetting on dielectric (EWOD) ielectric breakdown ontact angle hysteresis ydrophobicity ilanization\na b s t r a c t\nElectrowetting is a phenomenon that has received increasing attention in recent years. It offers much promise for applications in lab-on-a-chip devices, liquid lenses, and optical displays. The contact angle can be controlled electrically facilitating the transport of liquid droplets by electrocapillary forces. This paper focuses on the requirements of the dielectric layers in electrowetting on dielectric; firstly, to provide the required dielectric qualities, that is, sufficient dielectric strength and high capacitance and, secondly, to satisfy the surface wetting characteristics, that is, sufficiently high hydrophobicity and low contact angle hysteresis. We present experimental work that, firstly, compares the dielectric characteristics of several insulating layers and secondly, proposes silanization as a way to form hydrophobic layers for electrowetting. The relation between electrowetting and dielectric breakdown was tested by means of measuring the contact angle of a water droplet placed between two electrodes on an insulated surface. The leakage current and resistance between these electrodes were measured as a function of applied voltage for the following dielectric coatings: silicon dioxide, silicon nitride and Parylene C. We compare the contact angle response of electrowetting on octadecyltrichlorosilane (OTS) monolayers and plasmaenhanced chemical vapour deposition Teflon-like fluoropolymer layers. The contact angle hysteresis of the\nlowe act lin OTS monolayers is clearly is sufficient to avoid cont\n. Introduction\nElectrowetting is an electrokinetic means of manipulating a roplet by electrical modification of its surface tension. It has xcited much interest because of its potential for application in labn-a-chip devices, liquid lenses, microfluidic systems and optical isplays [1\u20133]. An electrokinetic force pulls the droplet along the radient of the surface energy [4]. In this paper, we will concentrate n the electrical and surface wetting properties of the insulating ayers used in electrowetting on dielectric. Lippmann first described how an applied voltage leads to a hange in the contact angle [4]. In electrowetting on dielectric EWOD) the variation of the contact angle on application of a oltage U can be predicted by\nos (U) = cos (0) + \u03b50\u03b5 U2 (1)\n2d\nhere (0) is the contact angle of the liquid in absence of applied oltage, \u03b50 is the permittivity of free space, \u03b5 is the dielectric contant of the insulating layer, d is the thickness of the insulating layer\n\u2217 Corresponding author. E-mail address: Brian.Cahill@iba-heiligenstadt.de (B.P. Cahill).\n925-4005/$ \u2013 see front matter \u00a9 2009 Elsevier B.V. All rights reserved. oi:10.1016/j.snb.2008.12.041\nr and this results in reversible electrowetting, that is, the hydrophobicity e pinning and dewetting occurs on removal of an applied potential.\n\u00a9 2009 Elsevier B.V. All rights reserved.\nand is the liquid surface tension. Contact angle saturation is a phenomenon that limits the validity of Eq. (1) [5].\nIn this paper, we measure the contact angle, leakage current and resistance of an electrowetting insulation layer with regard to applied voltage. We examine a number of dielectric coatings electrically in order to assess their viability for use in an electrowetting device. Moon et al. [6] were the first to focus on the dielectric layers used for EWOD with their main experimental setup being based on thermally oxidized silicon wafers. The relation between contact angle saturation and electrowetting was proposed by Papathanasiou and Boudouvis [7] and subsequently it was shown experimentally that contact angle saturation and dielectric breakdown occurred at the same applied voltages for very thin silicon dioxide layers but that dielectric breakdown can be improved with a stacked oxide\u2013nitride-oxide layer [8]. Unfortunately, thin silicon dioxide films suffer from significant problems with regard to dielectric breakdown and electrolysis when they are in contact with aqueous electrolyte [9\u201312]. Dash et al. [9] suggested that ion mobility in silicon dioxide films leads to an increased leakage current [13,14]. This paper will focus on the performance of silicon nitride and Parylene C as insulating layers in electrowetting.\nIn addition, we compare hydrophobization of silicon nitride by deposition of a plasma-enhanced chemical vapour deposition (PECVD) Teflon-like fluoropolymer layer with hydrophobization by", + "Actuators B 144 (2010) 380\u2013386 381\nm a l m t T u p m D a m u i B fl m p H s t\no F l t o e l t T c a l b t t t o o r s s m m c\n2\nu s l 1 a w t u p r t w w m o\nwith a 200 nm of PECVD-coated Teflon-like fluoropolymer layer. Although the Teflon-like fluoropolymer from INP e.V. displays a slightly higher contact angle, the layers on the samples from both suppliers showed static contact angles of greater than 90\u25e6.\nB.P. Cahill et al. / Sensors and\neans of a self-assembled monolayer of hydrophobic silanes. It is particular goal of our work, to find an alternative hydrophobic ayer to Teflon\u00ae AF. Most implementations of electrowetting have ade use of Teflon\u00ae AF, an amorphous fluoropolymer that is applied o the electrodes by spin-coating or dip-coating [1,6,8\u201310,15,16]. eflon\u00ae AF is without argument an excellent material choice for se in electrowetting. Nevertheless, its use has some significant ractical disadvantages: it is relatively expensive and, perhaps ore significantly, its use is limited by licensing agreements with uPont for both research and commercial purposes. Abdelgawad nd Wheeler [17] have opened the debate towards the developent of less expensive electrowetting devices and presented the se of SaranTM wrap (a consumer-grade polyethylene film) as an nsulating layer and Rain-x\u00ae water repellent as a hydrophobic layer. ayiati et al. [18,19] have claimed that PECVD deposited Teflon-like uoropolymer films offer similar hydrophobicity to that of comercially available spin-coated Teflon layers and also display low\nrotein adsorption with regard to biotechnological applications. ere, we examine the use of Teflon-like fluoropolymer film and a elf-assembled layer of silanes as hydrophobic layers in electroweting with regard to contact angle hysteresis.\nElectrowetting on top of self-assembled monolayers of thiols n gold electrodes has been reported by Sondag-Huethorst and okkink [20,21]. These monolayers were only stable within the imits of double-layer charging and at higher applied potentials hey provided no insulation to the flow of current and electrolysis ccurred. Both Lin et al. [11] and Saeki et al. [12] performed some xperimental work on the suitability of silanes as a hydrophobic ayer for electrowetting. To this end, they assembled octadecylrichlorosilane (OTS) monolayers on silicon dioxide thin films. hese setups were limited by the onset of electrolysis presumably aused by the dielectric breakdown of the silicon dioxide layer used s electrical insulation. Mugele et al. [22,23] utilized silane monoayers bound to oxide surfaces without focussing on their use. In oth cases, breakdown may have been avoided by using relatively hick insulating layers: a 50- m thick glass cover slip and a 1- m hick thermally grown silicon dioxide layer. Silanes bind covalently o surface hydroxyl groups; and silicon dioxide is by far the most bvious surface to provide such conditions. It may be that the dearth f research into electrowetting on top of hydrophobic silanes is a esult of the low resistance to electric breakdown exhibited by thin ilicon dioxide films in EWOD [11,12]. In this paper, we have choen to examine the performance of OTS because it is one of the ost widely used and cheapest of the available silanes. Furtherore, we bind OTS monolayers to a silicon nitride surface and as a onsequence avoid the use of a silicon dioxide insulating layer.\n. Methods\nFigs. 1 and 2 show schematic drawings of the setup that was sed to measure the electrical and surface wetting properties of a essile droplet sitting on top of an electrode structure that is insuated with a thin-film dielectric coating. For each experiment, a - L droplet was dispensed by means of a syringe pump in the rea between two parallel electrodes. A direct current (DC) voltage as applied between these electrodes using a Keithley 6517 elecrometer. The electrometer is capable of generating DC voltages of p to \u00b11 kV and of measuring leakage currents down to femtoameres (fA). The electrometer measured the current flow and the esistance of the system as a function of applied voltage. The elecrometer is controlled via the general purpose interface bus (GPIB), hich also enables digital transfer of measurement data. The setup as placed in a metal Faraday cage in order to minimise electroagnetic disturbances. It took between 2 and 5 s to measure each\nf the data points in the experiments presented here.\nParallel electrodes were used for the electric breakdown experiments. The droplet was placed between the electrodes and electrical measurements were performed between these electrodes as described in the above paragraph. The following coatings were applied: (a) a platinum electrode coated with 400 nm of Si3N4 and with a further 100 nm of Teflon-like fluoropolymer, (b) a copper electrode coated with 1.2 m Parylene C and with a further 200 nm of Teflon-like fluoropolymer, and (c) a gold electrode with 800 nm of SiO2 and 200 nm of Teflon-like fluoropolymer. Parylene C was deposited by vapour phase deposition. Parylene C and silicon nitride samples were hydrophobically coated as follows: the silicon nitride sample was coated with a further 100 nm of PECVD-coated Teflon-like fluoropolymer by GeSiM mbH (Gro\u00dferkmannsdorf, Germany) and the Parylene C sample was coated by the Leibniz Institute for Plasma Science and Technology (INP) e.V. (Greifswald, Germany)", + "3 Actuators B 144 (2010) 380\u2013386\no n i v c t s c p d\na w d w c o h s w s 2 t v s T c m g [\n3\nc a f 1 l a\n3\ns m s w h r A\nr t o a w t a l s c a c\nImages of the contact angle variation were acquired by means f a video camera equipped with a zoom lens. The synchroization output of the electrometer was connected to the sound\nnput of the audio/video adapter of the computer. The droplet ariations were recorded as video files by connecting the video amera to the digital video input of the audio/video adapter of he same computer. This allowed contact angle variation to be ynchronized with the electrical measurements. Additional static ontact angle and contact angle hysteresis measurements were erformed using a DataPhysics OCA 20 contact angle measuring evice.\nOTS, toluene, anhydrous sodium sulphate, ethanol, and nitric cid (65%) were purchased from Sigma\u2013Aldrich and were used ithout further purification. Toluene was dried by addition of anhyrous sodium sulphate prior to use. Silicon wafers were metallized ith a 100-nm layer of titanium platinum and were subsequently oated by PECVD silicon nitride. The silicon nitride surfaces were xidized by immersion in nitric acid (65%) for 1 min. This treatment as been used in the oxidization of silicon nitride AFM tips prior to ilanization [24]. Afterwards the samples were rinsed in deionized ater and then ethanol. Subsequently, the chips were dried in a tream of ultra-clean compressed air and then heated for at least h at 80 \u25e6C. After that, the samples were immersed in an octadecylrichlorosilane solution (1% OTS in dry toluene) for at least 12 h. The essel containing the OTS solutions was kept under an argon atmophere and in the dark to prevent bulk polymerization of the OTS. he samples were then rinsed with ethanol and dried with ultralean compressed air. OTS binds covalently to hydroxl groups by eans of its trichlorosilane headgroup. The trichlorosilane headroup acts to impose a well-ordered structure on the monolayer 25].\n. Results and discussion\nIt is of great importance for the viability of an electrowetting hip that the dielectric insulator should be (a) highly insulating nd (b) have a hydrophobic surface. In practice, this means the surace should have a static contact angle that is at least greater than 00\u25e6. In a first step, the breakdown characteristics of various insuating layers are compared and, in a second step, OTS monolayers re presented as a suitable hydrophobic layer for electrowetting.\n.1. DC breakdown characteristics of insulating layers\nThe electrical and surface wetting characteristics of silicon oxide, ilicon nitride and Parylene C coated with Teflon-like fluoropolyer were studied with regard to an applied voltage. In order to tudy electrical breakdown of insulating layers in electrowetting, e chose to use a PECVD-coated Teflon-like fluoropolymer as the ydrophobic layer. PECVD-coated Teflon-like fluoropolymer has ecently been proposed by Bayiati et al. as an alternative to Teflon\u00ae\nF. Fig. 3 shows the contact angles on the left-hand side, L and\night-hand side, R, of a droplet positioned between platinum elecrodes coated with 400 nm of Si3N4 and with a further 100 nm layer f Teflon-like fluoropolymer. The voltage dependence of the contact ngle on either side of the droplet is in relatively good agreement ith a parabola taking the form of Eq. (1) until, at about 80 V, conact angle saturation begins to become apparent. That is, the contact ngle stops changing with increasing applied voltage and Eq. (1) no onger predicts contact angle. It can also be seen that L and R aturate to slightly different values. The pH dependence of surface harging results in a negative potential at hydrophobic surfaces [26] nd Quinn et al. [16] showed how this causes some differences in ontact angle saturation on applying negative or positive bias volt-\nages. It is likely that this causes the observed difference in L and R.\nAs shown in Fig. 4 for a platinum electrode coated with 400 nm of silicon nitride and with a further 100 nm of Teflon-like fluoropolymer, the voltage was increased in discrete steps from 0 V to a maximum voltage. In three individual sweeps, the maximum voltage was increased consecutively. In the first two sweeps, to 200 and 220 V respectively, the leakage current remained below 10 nA although current increases and resistance decreases significantly at voltages above 200 V. In the subsequent voltage sweep, it is clear from the beginning of the sweep that the resistance of the dielectric layer has decreased significantly.\nIn the last sweep, the resistance was at a minimum half that of the first two sweeps and breakdown is also clear from 150 V. From this it can be concluded that it is very likely that dielectric breakdown occurred after the second sweep. This conclusion is backed up by the rapid increase in the current above 150 V during the third voltage sweep. Partial damage to the coating may have reduced the voltage at which current began to flow on following sweeps. A comparison of Figs. 3 and 4 shows that that contact angle saturation occurs at around 100 V while breakdown first occurs at around 180 V. This is not in agreement with the suggestion of Papathanasiou et al. [7,8] that contact angle saturation is caused by electric breakdown. It is very likely that for insulating layers of lower thickness, for example, 130 nm of silicon dioxide as used by Papathanasiou et al. [8], that contact angle saturation and dielectric breakdown commence at the same applied potential.\nFig. 5 displays two voltage sweeps performed on a copper electrode coated with 1.2 m Parylene C and with a further 200 nm of Teflon-like fluoropolymer. In the first sweep, the maximum applied voltage was 500 V, no breakdown is apparent in either the leakage current resistance characteristic curves. In the second voltage sweep, breakdown was apparent at a voltage slightly above 850 V.\nPreliminary results with plasma-coated silicon dioxide proved to be extremely disappointing and will not be shown graphically. In general, much higher currents and much lower resistances were observed before breakdown and breakdown itself occurred at quite low voltages. As discussed in Section 1, these results are similar to the difficulties with dielectric breakdown of silicon oxide in electrowetting experiments that others have reported [7,9\u201312]." + ] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure13.4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure13.4-1.png", + "caption": "Fig. 13.4 The Jazzy 1122", + "texts": [ + "3, it is clear that the sensor does not return a value linear or proportional to the actual distance because the intensity of the infrared signal is inversely 30313 Web-Based Control of Mobile Manipulation Platforms via Sensor Fusion BookID 175907_ChapID 13_Proof# 1 - 12/07/2009 probational to the square of the distance. Therefore, the infrared signal falls rapidly as the distance increases. As described in (Jazzy 1122 the owner\u2019s manual [12]), the jazzy wheelchair has two main assemblies: the seat and the power base as described in Fig. 13.4. Typically, the seating assembly includes the armrests, seatback, and controller. The power base assembly includes two drive wheels, two anti-tip wheels, two rear caster wheels, and a body shroud. In our project, we remove the armrests and seatback as shown in Fig. 13.5. The specifications of the Jazzy 1122 wheelchair are described in Table 13.1. The jazzy 1. Active-Trac Suspension: The wheelchair is equipped with Active-Trac Suspension (ATS) to be able to traverse different types of terrain and obstacles while maintaining smooth operation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002109_physrevlett.107.268101-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002109_physrevlett.107.268101-Figure3-1.png", + "caption": "FIG. 3. The resistive force and torque on a cylinder rotating with speed ! and translating with velocity Vfx\u0302 in a fluid between two concentric cylinders. (a) The cell cylinder and the outer membrane are considered to be concentric cylinders with a fluid in between them. The cell cylinder has radius R, and the outer membrane has radius R\u00fe . A single cylinder of radius a rotates in the space between these two cylinders. (b) The geometric parameters for the region about the flagellum.", + "texts": [ + " It is important to note that the resistive force and torque for this scenario are determined by a single drag coefficient. Next, we consider the case where the flagellum is not in direct contact with the cell cylinder, and, instead, a layer of fluidmediates the interaction between the flagellum and the cell cylinder and outer membrane. The periplasmic space is treated as the region between two concentric cylinders of radii R and R\u00fe , where the inner cylinder is the cell cylinder and the outer cylinder represents the outer membrane [Fig. 3(a)]. The periplasm is filled with fluid of viscosity p. Rotation of the flagellum in this space is opposed by fluid drag caused by the relative motion of the flagellum with respect to the cell cylinder and the outer membrane. In Borrelia, the distance from the flagellum to the cell cylinder, h , and the distance from the flagellum to the outer membrane, h\u00fe, are roughly comparable to the radius of the flagellum. The case of a cylinder rotating and translating near a single wall (e.g., h\u00fe ! 1) has been solved analytically [12]", + " We have analyzed this problem using lubrication theory and numerical analysis [14,15] and have found that the force on the rotating cylinder arises from the constraint that the volume flux of fluid in the gap above the cylinder must equal the flux of fluid through the lower gap. This simple constraint can be used to derive approximate relationships between the resistive force, torque, rotational speed, and flagellar velocity [16]. We first note that rotation of the outer membrane with respect to the cell cylinder in Fig. 3(a) will produce a resistive torque on the cell cylinder that is approximately equal to 2 pR 2\u00f0V\u00fe V \u00de= [17]. In addition, rotation of the flagellum can produce torques on the cell cylinder and outer membrane. Since is comparable in size to a, the magnitudes of these torques scale independently of . Therefore, to leading order in , the outer membrane and the cell cylinder rotate together, and V\u00fe V [Fig. 3(b)]. We can then set Vf to be the velocity of the flagellum relative to the cell cylinder and outer membrane. For B. burgdorferi, the flagellar diameter is roughly half of the width of the periplasmic space. Therefore, in what follows we assume that \u00bc 4a. It is also useful to define the ratio of the top and bottom gaps, h \u00bc h\u00fe=h . The resistive force and torque on the flagellum are then approximately [14] fr \u00bc 6 ffiffiffi 2 p p\u00f01\u00fe h\u00de5=2 \u00f01\u00fe h5=2\u00de 1 h 1\u00fe h a! 5Vf (2) r\u00bc 2 a p\u00f01\u00feh\u00de5=2 \u00f01\u00feh5=2\u00de 5\u00f01 h\u00de 2\u00f01\u00feh\u00deVf \u00f0h 1\u00feh 1\u00dea" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003211_s12541-013-0227-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003211_s12541-013-0227-3-Figure1-1.png", + "caption": "Fig. 1 Experimental setup", + "texts": [ + " However, the micro endmills in their studies were used in a profile side cutting configuration, and not fully immersed in the workpiece. In this research, laser-assisted mechanical micromilling experiments have been performed for biomedical implant materials: titanium alloy Ti6Al4V, stainless steels AISI 316L (SS 316L) and 422 (SS 422). Thermal and mechanical modeling analyses are conducted to analyze the performance of LAMM for these difficult-to-machine implant materials. Cutting temperatures and workpiece flow stresses are characterized in the cutting zone. The laser-assisted micro milling experimental setup is shown in Fig. 1. The LAMM experiments were carried out on a CNC controlled system with 1 \u00b5m resolution three-axis stages, which is equipped with a high speed spindle (Precise SC-40) with a maximum rotation speed of 90 k RPM. A 25 W CO2 laser (Synrad 48-2) and optics were used in a sturdy experiment bench. A flexible nozzle was attached to the spindle mounting fixture allowing for an adjustable flow of assist gas during LAMM. Two-flute, 100 \u00b5m diameter tungsten carbide endmills were used. Experiments were performed on rectangular workpieces of approximately 25(L) \u00d7 20(W) \u00d7 15(H) mm in size" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003575_iros.2011.6094760-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003575_iros.2011.6094760-Figure1-1.png", + "caption": "Fig. 1. Diagram of a UAV at position ~rA that is moving at heading angle \u03b8 and tracking a randomly-moving target at ~rT with distance r = |~rA \u2212 ~rT | and relative angle \u03d5.", + "texts": [ + " In Section IV we analyze the connection between the chosen noise intensity and the resulting control law, as well as demonstrate the effectiveness of this approach for both Brownian targets and targets with unknown trajectory. Section V concludes this paper and provides directions for future research. We consider a UAV flying at a constant altitude in the vicinity of a ground-based target, tasked with maintaining a nominal distance from the target. The target is located at position ~rT (t) = [xT (t), yT (t)] T at the time point t (see Fig. 1). The UAV, located at position ~rA(t) = [xA(t), yA(t)] T , moves in the direction of its heading angle \u03b8 at a constant speed vA. The turning rate is determined by a non-anticipative [11], bounded control u(t) \u2208 U \u2261 {u : |u| \u2264 umax}, which has to be found. In our problem formulation, the target motion is unknown. We therefore assume that it is random and described by a 2D stochastic process. Drawing from the field of estimation, the simplest signal that can be used to describe an unknown model suggests that the motion of the target should be described by a 2D Brownian particle: dxT (t) = \u03c3dwx dyT (t) = \u03c3dwy (1) where dwx and dwy are increments of unit intensity Wiener processes along the x and y axes, respectively, which are mutually independent", + " The UAV can be modeled as a planar Dubins vehicle [1]: dxA(t) = vA cos (\u03b8(t)) dt dyA(t) = vA sin (\u03b8(t)) dt d\u03b8(t) = \u2212u(t)dt, u \u2208 U (2) where, without loss of generality, we have chosen the sign of d\u03b8(t) to clarify later results. In order for the control to be independent of the heading angle of the UAV or the absolute position of the UAV or target, we relate the problem to relative dynamics based on a time-varying coordinate system aligned with the direction of the UAV velocity. The reduced system state is composed of the distance between the UAV and target r = |~rT \u2212 ~rA| and the viewing angle \u03d5 between the UAV\u2019s direction of motion and the vector from the UAV to the target, as seen in Fig. 1: r = \u221a (\u2206x)2 + (\u2206y)2 (3) \u03d5 = tan\u22121 ( \u2206y \u2206x ) . (4) The combined UAV-target system (1-4) should maintain the relative distance r at the nominal distance d for all times. To this end we seek to minimize the expectation of an infinite-horizon cost function W (\u00b7) with a discounting factor \u03b2 > 0 and with zero penalty for control: W (r, \u03d5, u) = Eu r { \u222b \u221e 0 e\u2212\u03b2tk (r(t)) dt } (5) k(r) = (r \u2212 d) 2 . A high value of \u03b2 places more weight on the instantaneous cost, while \u03b2 near 0 considers future costs, as well as instantaneous cost" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003602_tmag.1970.1066758-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003602_tmag.1970.1066758-Figure2-1.png", + "caption": "Fig. 2 Arrangement of permanent magnets in a) 2-pole system, b) l&pole system.", + "texts": [ + " EXPERIMENTAL ARRANGEMENT Fig. 1 shows the buildup of the braking system. There are two iron disks fitted with barium ferrite 300. The distance between the two disks is variable. Between the two disks, there is a d = 5 mm thick eddy-current disk of copper or aluminum which was driven at different speeds. The forces exerted on the fixed system of permanent magnets can easily be measured with the aid of a spring scale. The investigation covered magnet systems with 2, 4, 6, 8, 10, 12, 16, and 18 poles. Fig. 2 shows how the permanent magnets are arranged on the return disk in the case of the 2-pole and the 18-pole systems. A.fter removal of the eddy-current disk, the field strength at its site was measured with a Hall-effect instrument for different distances 6 between the permanent magnets. The field profiles relating to the individual 6 distances differ only by a numerical factor. Fig. 3 shows the field profiles, which were measured in azimuthal direction and normed to I, for the 2-pole and the 18-pole magnet system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003692_978-3-642-23681-5_13-Figure13.5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003692_978-3-642-23681-5_13-Figure13.5-1.png", + "caption": "Fig. 13.5 Dynamic test rig layout (Cabrera et al. [28])", + "texts": [ + " The measurements of pressure indicated that the film pressure profiles are very different from those of the conventional rigid bearings. The relatively low pressures in the film caused significant rubber deflections but too low to produce viscosity changes. Integration of the pressure in the bearings showed that they operate in the regime of mixed lubrication. The behaviour of the bearings was theoretically investigated using CFD and compared with the experimental values conducted on the test rig designed (See Fig. 13.5). The bearing test rig was used for 50 mm diameter bearings and lengthto-diameter ratio of 2. The test shaft was made of chrome-plated stainless steel and supported by two hybrid bearings. The test loads were applied using a motorised unit, a de-coupling spring and a load cell. The loading unit allowed loads of up to 1,000 N to be applied. The tests were carried out with a shaft surface speed of 2 m/s and loads in the range 0\u2013500 N. Production bearings manufactured by Silvertown UK Ltd. were used for this investigation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002790_s11465-013-0254-x-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002790_s11465-013-0254-x-Figure1-1.png", + "caption": "Fig. 1 Schematic of the cam mechanism with oscillating roller follower", + "texts": [ + " In this paper a theoretical formulation to describe the nonlinear dynamic behaviour of a cam mechanism with oscillating roller follower is presented. A lumped parameters model with eight degrees of freedom is developed. The nonlinear dynamic behaviour is described by a system of non linear second order differential equations. The dynamic responses are obtained through a coupling between the implicit Newmark algorithm with the iterative Newton-Raphson method to iterate the solution for each time step. Figure 1 shows the cam mechanism with oscillating roller follower studied in this paper. This mechanism consists of a cam (1) mounted on a camshaft, an oscillating follower (3) with 2 rollers (2) and (4) and a sliding rod (5) that translate vertically. A spring is inserted between the sliding rod (5) and the frame to maintain two contacts at points C1 and C2. Cam (1) rotates at a constant angular velocity \u03c91. This rotation causes the oscillation of the follower (3) relatively to a fixed point O3 and hence the translation of the sliding rod (5)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001413_20090916-3-br-3001.0009-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001413_20090916-3-br-3001.0009-Figure2-1.png", + "caption": "Fig. 2. Coordinate systems.", + "texts": [ + " In the fourth section SM controller applied to DP system is described. Five and six sections describe the tuning of controller parameters and experimental set-up, respectively, while the seventh section shows the results of the experiments. Finally, the conclusions are present in the eighth section. Dynamic Positioning Systems are only concerned with the low-frequency horizontal motions of the vessel, that is, surge, sway and yaw. The motions of the vessels are expressed in two separate coordinate systems (see Fig. 2): one is the inertial system fixed to the Earth, OXYZ; and the other, O\u2019x1x2x6, is a vessel-fixed non-inertial reference frame. The origin for this system is the intersection of the midship section with the ship\u2019s longitudinal plane of symmetry. The axes for this system coincide with the principal axes of inertia of the vessel with respect to the origin. The motions along of 978-3-902661-51-7/09/$20.00 \u00a9 2009 IFAC 237 10.3182/20090916-3-BR-3001.0009 the axes O\u2019x1, O\u2019x2 e O\u2019x6 are call of surge, sway e yaw, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003618_s11431-013-5291-5-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003618_s11431-013-5291-5-Figure4-1.png", + "caption": "Figure 4 Numerical model of inner ring containing elliptical cracks: (a) Whole numerical model; (b) subsurface crack definition.", + "texts": [ + " Subsurface cracks are modeled as three elliptical pores and located at an approximately depth H (0.2 mm) below the raceway surface. The major size 2a, the minor size 2b and the thickness 2c of these cracks are 0.5, 0.25 and 0.25 mm, respectively, as shown in Figure 3. Symbol \u03b8 indicates the rotation angle of the inner ring. The rotation angle is defined to be positive when the contact position is to the right of the cracks. The moment M is exerted on the inner ring in the anticlockwise direction. The numerical model is developed as shown in Figure 4(a). The material used in this study is AISI 52100 steel and the main parameters are: Young\u2019s modulus E=2.10\u00d7105 MPa, Poisson\u2019s ratio =0.3, yield stress y=1747 MPa, and failure stress r=2106 MPa. The inner ring was meshed with 20-noded quadratic hexahedron elements. To study the crack propagation, it is necessary to define the crack front, crack tip and crack extension direction [23], as shown in Figure 4(b). The elliptical surface is defined as the crack front, the edge is defined as the crack tip, the direction parallel to the major size is assumed as the crack extension direction. Using this method, three subsurface cracks, distributed along the \u201cz\u201d direction, are defined, as shown in Figure 5, and their interval s is varied to investigate RCF in the bearing ring. The three crack tips are called tips 1, 2 and 3. To get accurate stress and strain, the refined mesh is created around the subsurface cracks" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003889_isie.2013.6563767-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003889_isie.2013.6563767-Figure9-1.png", + "caption": "Fig. 9", + "texts": [ + " Fuzzy algorithm is easy to program and not complex. Data are collected from five directions (see Fig. 7); each datum represents the distance between the WMR and obstacles. Fig. 8 shows the condition when the WMR making a turn. There are two wall following method. One is \u201cleft wall following\u201d and the other is \u201cright wall following\u201d. For choosing which one method is depend on which the direction of datum is closer to the obstacle. The Input data are front distance D3 and \u221a or \u221a . Output is the turn-angle. Fig. 9 and Fig.10 show the membership function of input. Fig. 11 shows the membership function of output. Fig. 12 and Fig.13 show the fuzzy rules of the two wall following. These rules can be described below: if S1 is Negative and S2 is Near , then y is Right\u3002 if S1 is Negative and S2 is Far , then y is Left\u3002 . . . if S1 is Positive and S2 is Far, then y is Middle\u3002 V. SIMULATION AND EXPERIMENT 1) Simulation 1 : In Fig.14, the red line shows the whole path from start point to the goal. When encountering an obstacle in front of the WMR, the WMR will choose a path of the upside and downside, randomly" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001657_j.jbiomech.2008.10.040-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001657_j.jbiomech.2008.10.040-Figure3-1.png", + "caption": "Fig. 3. The geometry of the active (left) and passive cilium (right). Both of the cilia are modeled with a cylindrical body (5.8mm in height, 0.2mm in diameter) and a parabolic tip (0.2mm in height). The active cilium is bended with a curvature of p/ 6, and the distance between the active and passive cilium is 9.35mm.", + "texts": [ + " According to experimental observations (Nonaka et al., 2005), the sheath of the cilia is modeled as a composite of a cylindrical body (5.8mm in height, 0.2mm in diameter) and a parabolic tip (0.2mm in height). The initial position of the passive cilium is straight while the active one is bent with p/6 curvature towards the positive x-direction. The spatial resolution of the cilia involves 107 points in the vertical direction (100 points of their cylindrical body and 7 points of the parabolic tip) and 16 points around the circumference (Fig. 3). Both the cylindrical volume (for the active cilium), and the cuboid domain (for the passive one), were discretized using an unstructured tetrahedral grid. The basic grid configuration consists of a total of 365,554 cells (grid independency is discussed in the sequel). The fluid domain is considered aqueous (Buceta et al., 2005). The stiffness of the passive cilium has not been reported directly so far. In Okuno et al. (1981) the stiffness of the flagella, S. purpuraturs spermatozoa, was reported to be 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000233_tie.2007.898297-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000233_tie.2007.898297-Figure3-1.png", + "caption": "Fig. 3. MEMS-based pneumatic microactuator principles. (a) Front-side view. (b) Back-side view.", + "texts": [ + " Several different hardware architectures were tested to make the FPGA resource amount as low as possible, and the controller/manager as simple and efficient as possible. In Section II, a brief approach of the MEMS chip fabrication process is given. Section III describes the control architecture applied to the distributed micromanipulator. The decentralized decision-making strategy, developed in very high-speed hardware design language (VHDL) and implemented in an FPGA target, is detailed. Section IV presents integration and experimental results of the distributed control approach. Section V is the conclusion. Fig. 3 describes principles of the MEMS-based distributed pneumatic microactuators we developed for air-flow planar micromanipulator. Illustrations give views of distributed airflow surface with object, and focus on front-side and back-side of the single pneumatic microactuator. At the bottom of the substrate, a mobile microvalve is supported by suspension beams and actuated by electrostatic effect. Two electrodes generate electrostatic force are aligned next to the mobile microstructure part and fixed to the top substrate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001265_ichr.2009.5379602-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001265_ichr.2009.5379602-Figure3-1.png", + "caption": "Fig. 3. Mapping space and frame coordinates. The tilting angle of the stereo vision in z-axis is about 10 degrees.", + "texts": [ + " Thus, our approach to the second question is that the grid matching is conducted with the mean of all the For the moment, our purpose on humanoid SLAM is not the object recognition level but the path planning level. Therefore, the resolution of a grid or a voxel is set to IOcm in this work. For the generation of a grid map, the same method to [4] is used, but in our work, real images taken from HRP-2 are used. The image size of HRP-2 is 640x480, and the ranges of an image in the z-axis are from O.5m to 4m. Note that only depth value of a pixel is used to build a grid map but neither width nor height value is used. The size of mapping space as shown in Fig. 3 is 2m in width and height, and 3m in depth. In the figure, the coordinate notations for the global frame, the camera frame, and the image frame are denoted by Ea, Ee, El , respectively. The procedure to generate a grid map from range data of stereo images is summarized in Table I, based on the coordinate frames. In a nutshell, first, a voxel in the mapping space is projected to the image frame, denoted by Pg (line 2 in Table I). Second, the occupancy probability of the voxel is computed based on the difference between the depth of the voxel and the depth in the projected pixel in the image, which are denoted by Dg and Dp , respectively (line 7-10) " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001315_978-3-540-30301-5_27-Figure26.2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001315_978-3-540-30301-5_27-Figure26.2-1.png", + "caption": "Fig. 26.2 A mobile manipulator inserting a peg into a hole", + "texts": [ + " The end-effector is so named as it is often the end link in a serial kinematic chain. However, complex manipulation tasks may require the simultaneous exertion of multiple forces and moments at different contact points as well as the execution of sequences of force applications to several objects. To understand the challenge of generating the motion required for the execution of a manipulation task, we will consider the classic example of inserting a peg into a hole. The mobile manipulator shown in Fig. 26.2 is commanded to perform the peg insertion manipulation task, whose general problem structure arises in a number of contexts such as compliant assembly. For simplicity, we will only consider the transfer motion (see Sect. 26.3) and assume that the robot has already established a stable grasp (see Chap. 28) and is holding the peg. If the clearance between the peg and the hole is very small, the peg may not be able to enter the hole easily, and often some contact occurs. In order to guide the peg successfully into the hole, the robot has to deal with the possible contact states between the peg and the hole (Sect", + " Through contact points the robot is able to apply forces and moments to objects. By controlling the position and velocity of the contact points as well as the forces acting at them, the robot causes the desired motion of objects, thereby performing the manipulation task. The programming of such a task is most conveniently accomplished by directly specifying positions, velocities, and forces at the contact points, rather than by specifying the joint positions and velocities required to achieve them. Consider the manipulation task of serving a cup of water shown in Fig. 26.2. This task can easily be specified by providing a trajectory for the cup\u2019s motion. To determine a joint space trajectory that achieves the same motion, however, would be much more complex. One would have to rely on inverse kinematics (Sect. 1.9), which can be computationally challenging. More importantly, the task constraints imposed on the cup\u2019s motion do not uniquely specify the cup\u2019s trajectory. For example, while the cup is not allowed to tilt, it can be delivered to the goal location in any vertical orientation", + "5), the corresponding joint torques used to control the robot are computed based on (26.1). The full expressiveness of task-level control comes to bear in the context of redundant manipulators. A manipulator is considered redundant with respect to the task it is performing if has more degrees of freedom than required by the task. For example, the task of holding a cup of water only specifies two degrees of freedom, namely the two rotations about the axes spanning the horizontal plane \u2013 this task has two degrees of freedom. The mobile manipulator shown in Fig. 26.2 has ten degrees of freedom, leaving eight redundant degrees of freedom with respect to the task. 2 6 .2 The operational space framework for task-level control of redundant manipulators decomposes the overall motion behavior into two components. The first component is given by the task, specified in terms of forces and moments, Ftask, acting at an operational point. This vector F is translated into a joint torque based on (26.2): \u03c4 = J Ftask. For a redundant manipulator, however, the torque vector \u03c4 is not uniquely specified and we can select from a set of task-consistent torque vectors", + " The main difficulty of assembly motion is due to the requirement for high precision or low tolerance between the parts in an assembled state. As a result, the assembly motion has to overcome uncertainty to be successful. Compliant motion is defined as motion constrained by the contact between the held part and another part in the environment. As it reduces uncertainty through reducing the degrees of freedom of the held part, compliant motion is desirable in assembly. Consider the peg-in-hole insertion example introduced earlier (see Fig. 26.2). If the clearance between the peg and the hole is very small, the effect of uncertainty will most likely cause a downward insertion 2 6 .4 motion of the peg to fail, i. e., the peg ends up colliding with the entrance of the hole in some way without reaching the desired assembled state. Therefore, a successful assembly motion has to move the peg out of such an unintended contact situation and lead it to reach the desired assembled state eventually. To make this transition, compliant motion is preferred" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002736_tmag.2012.2196502-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002736_tmag.2012.2196502-Figure7-1.png", + "caption": "Fig. 7. Distributions of residual stresses in R.D. and T.D.", + "texts": [ + " The distributions were similar to the distributions because the flux condition was almost alternating in the ring-core model. Since the tendency of the magnetic characteristic distribution depending on the residual stress agrees with the measured one (Fig. 2), it can be said that the SCES modeling is applicable to magnetic characteristic analysis considering stress effect in core materials. The developed method was applied to analysis a PM motor model core as shown in Fig. 6. Table II shows the conditions used in the analysis. The motor model core has 4 poles and 36 slots. Fig. 7 shows the distributions of the residual stress in R.D. and T.D. The residual stress in the stator core was measured by using the X-ray stress measurement system [4]. The rolling direction agrees with the X-direction. However the local easy axis depends on the residual stress distribution. The compressive stress in R.D. and T.D. was distributed mainly in the back yoke and the largest tensile and compressive stress occurred around the caulking. Therefore, it can be considered that the local magnetic properties in these areas are deteriorated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002160_2011-01-1548-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002160_2011-01-1548-Figure1-1.png", + "caption": "Figure 1. Schematic of a 14-DOF nonlinear dynamic model of hypoid geared rotor system.", + "texts": [ + " On the other hand, the dynamic bearing load is found to be affected significantly by the temporal variation in its stiffness, as expected. The jump response frequency and amplitude of both the dynamic mesh force and dynamic bearing load are affected only slightly by the time-varying bearing stiffness property. On the other hand, both the dynamic mesh force and bearing loads are found to be sensitive to backlash. A generic hypoid geared rotor system consisting of a coupling gear pair, shafts, bearings, engine and load is modeled with a nonlinear, 14 degrees of freedom (DOF) lumped parameter representation as shown in Figure 1. Pinion and gear are considered as rigid bodies, and the engine and load are each represented by a rotational coordinate. The flexibility of both pinion and gear shaft bearing assemblies is condensed into a set of stiffness matrices to support the pinion and gear bodies. The coupling between the engaging gear pair is represented using single-point mesh model. The proposed mesh model consists of mesh point, line-of-action, mesh stiffness, backlash and transmission error. The model used in this paper is similar to the one proposed by Tao[17] except for the following feature", + " The ordinary differential equations of motion of the system can be written as (10) (11) (12) where \u03b8E and \u03b8L stand for the torsional coordinates of engine and load, and (xl,yl,zl,\u03b8lx,\u03b8ly,\u03b8lz) stand for the translational and rotational coordinates of the pinion (l = p) or gear (l = g). Correspondingly, IE and IL are the inertia masses of engine and load, and (Ml,Ilx,Ily,Ilz) are the masses and inertia masses of the pinion (l = p) or gear (l = g). Note that the coordinates of pinion and gear are represented in pinion coordinated system (Sp) and gear coordinate system (Sg) as shown in Figure 1. The stiffness matrix [K] consists of the contributions from condensed shaft and bearing support stiffness matrices of both pinion and gear, and it is defined as follows: (13) where the two constants kE and kL stand for the torsional stiffnesses of the input and output shafts. The value of system damping matrix [C] depends on the type of damping model employed; for example, modal damping, system damping and component damping models. The force vector {F} in equation (10) includes the external excitation and internal forces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002924_mcs.2012.2234971-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002924_mcs.2012.2234971-Figure2-1.png", + "caption": "Figure 2 Basic maneuvers versus rotor speeds. Faster and slower rotors are shown by bigger and smaller circles, respectively. roll and pitch movements are obtained by increasing and decreasing opposite (4\u20132 and 1\u20133) rotor speeds. This figure is an approximation since the coupled dynamics necessitate trimming every rotor speed to maintain horizontal stability. Additionally, the total vertical thrust must be kept constant to simulate a stabilized hovering condition regardless of maneuvers.", + "texts": [ + " The setup is fixed on a 3-DOF universal joint with !12\u00b0 freedom in the roll and pitch axes, and with !360\u00b0 freedom in the yaw axis. Actually, the yaw axis is limited by only the length of the cables attached to the fixture. Movement is not available on the z-axis. This limited configuration is ideal for attitude control since the dynamic equations are reduced by the elimination of unused axes that are x, y, and z [24]. The roll, pitch, and yaw movements are accomplished by varying the rotor speeds. This phenomenon is depicted in Figure 2. Bigger and smaller circles represent faster and slower rotors, respectively. The pitch and roll movements are obtained by increasing and decreasing opposite 66 IEEE CONTROL SYSTEMS MAGAZINE \u00bb April 2013 The Educational Environment The users of the setup interact mainly with the ControlDesk and Simulink programs. After powering up all of the nec- essary equipment, starting applications, and loading files, a demonstration or experiment is started within 5 min. This fea- ture is essential for in-class demonstrations since the time is limited" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001479_bf01231434-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001479_bf01231434-Figure2-1.png", + "caption": "Fig. 2. Positions of equilibrium at L4, m2> ml.", + "texts": [ + "10) showing that motion in this plane is periodic and that the period of infinitesimal oscillations about the origin is T0 = 4re(10 + 3 x / i + 12#2) -1/2 . (7.11) If ~, q, 4 and 0 are all regarded as first order small quantities the solutions of (7.5) and (7.6) have the forms = A sin (Pt + ~) = B cos (Qt + fl), where A and B are arbitrary first order small constants, e and fl are arbitrary constants and p 2 = _}2 x/1 + 12#2 = 2Q2 _ 5. Under these circumstances, the curves described are Lissajous figures; the periods in the 0 and 4~ directions being To and T 0 respectively. Figure 2 illustrates the positions of stable and unstable equilibrium at L4 (/*, x2x/3, 0). It can be seen that when the masses are equal, # - 0 and the stable position is parallel 330 w.J.ROBrNSON to OY. On the other hand, as mt >0, # ~ 89 and ~ >hi3, that is, the satellite points towards m2. For the Earth-Moon system ~=29 ~ 42' and To and T+ are of the order of 14 and 16 days respectively. Acknowledgements The author wishes to express his appreciation for many comments and suggestions made by Dr Leon Blitzer of the University of Arizona relevant to this paper" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure4-1.png", + "caption": "Figure 4: Two Roller Control Mechanism", + "texts": [], + "surrounding_texts": [ + "The increasing demand for improving vehicle fuel economy is guiding automakers into developing more efficient powertrain systems. One area receiving significant attention is the optimisation of the base engine itself where both downsizing and reducing the pumping losses of the engine are essential. Therefore, pressure charging the engine, either via a supercharger or a turbocharger, is becoming increasingly prevalent in both today\u2019s and in particular future automobiles. It is recognized that pressure charging a downsized engine provides increased performance and thermal efficiency. However it is also recognised that a downsized pressure charged engine lacks power and torque at lower engine speeds. In supercharger applications, in an attempt to alleviate this issue, the supercharger is normally \u201dgeared up\u201d with the advantage of providing sufficient boost at low engine speeds. The disadvantage of speeding up the supercharger are the losses incurred by spinning at high speeds and the requirement to bypass \u2013 at higher engine speeds the supercharger provides more boost than is needed and hence the system is bypassed introducing a further inefficiency into the system. An alternative to driving the supercharger directly would be to provide a variable drive \u2013 one that would speed the supercharger up at low engine speeds / when boosting is required and slow the supercharger down at higher engine speeds or when boosting is not required (highway cruising speeds). Hence, having determined that a variable speed drive to a supercharger is potentially beneficial, the next step was to determine the optimal arrangement including Variator ratio spread and supercharger speeds. Various gearing arrangements and Variator ratio spreads where investigated with the conclusion that a Variator ratio spread of 5 connected directly to the engine with a step up gearing on the output would provide the best arrangement of the system. To minimise package, the Torotrak Variator is normally arranged with a symmetrical variable ratio spread around the 1:1 point. Therefore, for a ratio spread of 5, a Variator with a maximum ratio of 2.236 (= \u221a5) and a minimum ratio of 0.447 (= 1/\u221a5) is applied. With this ratio spread, a step up gear of 2:1 between the Variator and the supercharger was assumed. This provides the potential envelope of supercharger operating speeds as shown in Figure 6. With a clear focus on achieving a low cost, weight and package solution, a variable drive for a supercharger needs to be based upon the two roller \u2018yoke\u2019 control Variator design from the OPE transmission. Two supercharger drive designs have been developed in single and twin cavity format with a high level of commonality between designs with both employing 50mm diameter rollers. The single cavity design is shown in Figure 7 : For a traction drive device to operate, the discs and rollers need to be clamped together. Various method of achieving the necessary clamping or \u201cEnd load\u201d force exist from complex hydraulic arrangements to simple bevel spring designs. Focussing on cost, the Variator geometry and associated gearing have been designed in conjunction with the speed and load characteristics of a supercharger to produce a relatively flat high reaction force End Load requirement curve (Figure 8). Therefore, application of a simple bevel spring arrangement to generate the clamping forces together with a thrust bearing to react the loads is possible. The significant advantage of the single cavity design is simplicity and low cost due to the low parts count. The compromise for this low-cost solution is efficiency \u2013 due to the end load being reacted through a thrust bearing, the single cavity Variator has an efficiency of ~82% to 83% at high reaction loads. Given this efficiency, a single cavity design utilising a 50mm roller diameter has a power capacity of ~ 15kW (constant). This is suitable for the lower cost, entry level supercharger systems. Regarding the axial loads, if a second toroidal cavity is introduced, the two cavities react the loads against each other so removing the need for the thrust bearings. This increases the efficiency to ~89% (at high reaction loads) and correspondingly, increases the power capacity \u2013 for the same 50mm roller diameter, ~45kW power capacity results. The compromise is increased size and parts count. However, as the thrust bearings have been deleted, the impact on length of the variable drive system is minimised. The twin cavity design is shown in Figure 9 : The twin cavity design therefore provides a power dense variable drive in a package compatible with automotive supercharger applications." + ] + }, + { + "image_filename": "designv11_3_0002138_1.4755980-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002138_1.4755980-Figure1-1.png", + "caption": "FIG. 1. Schematic illustration of the longitudinal section of the cut front in the feed direction.", + "texts": [ + " According to results of Hirano and Fabbro,2 this technique gives a rather accurate image of the current processing zone or the cut front shape, respectively, since the melt layer is so thin that it is frozen very rapidly when the laser is turned off. Experimental trials were not only conducted at the achievable maximum cutting speed but also for relative cutting speed values of v1\u00bc 0.33 and v2\u00bc 0.66 with respect to the maximum value Vmax. Longitudinal microsections of the cut front shape along the symmetry plane were prepared by sectioning the samples with a cut-off machine in cutting direction (Fig. 1). This cut was carried out parallel to the cut kerf in about 3 mm distance to the kerf edge. Subsequently, the specimens were embedded into transparent acrylic resin and ground with SiC paper (from 180 to 520 grit) just to the cut edge. After that, the remaining part of material was carefully removed by polishing the samples with 9 lm diamond paste. Images of the cut front shape in a digital data format were taken by use of the optical multifocus microscope Nikon AZ 100 M. Exemplarily, Figure 2 shows a rather smooth cut front geometry for the given parameter constellation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000718_j.triboint.2008.06.014-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000718_j.triboint.2008.06.014-Figure2-1.png", + "caption": "Fig. 2. Supply system schematics.", + "texts": [ + " This parameter depends on the diameter ds of the supply holes, on their arrangement if the holes are adjacent, on clearance h0 and on supply pressure ps. The dynamic behavior of bearings with different hole diameters can be compared if the hole downstream pressure level pc0 is the same. For this purpose, an analysis of the resistances at different supply pressures with the shaft in the central position was performed under both static and rotating conditions. A schematic view of the supply system is shown in Fig. 2. The Reynolds equation was solved for fixed geometry and with the initial condition p \u00bc pa at all points to find the steady-state solution. Multiple iterations were carried out in order to find the supply hole diameter which corresponds to a particular value of pc0 (1.5, 2, 2.5 and 3 times the ambient pressure), at ps \u00bc 0.3, 0.5 and 0.7 MPa. Fig. 3 shows a comparison of how pc0 varies with supply hole diameter ds, at different supply pressures for bearings with a single set of holes (Fig. 3a) and two sets of holes (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002203_0957-4565.41.10.76-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002203_0957-4565.41.10.76-Figure2-1.png", + "caption": "Figure 2. Position of the tested fault gears Figure 3. Typical gear faults", + "texts": [ + " A crack, wear or broken gear tooth failure may cause fatal accidents. Thus an early recognition of the gear tooth faults is critical for the regular operation of a gearbox. Experimental analysis of a two-stage gear transmission has been carried out using an experimental setup. The gearbox in our experiment is illustrated in Figure 1(left) and the fault simulator setup with sensors is shown in Figure 1(right). Gears #Z40 and #Z85 are the tested gears. The typical gear faults in our experiment are shown in Figure 2 and Figure 3. A variable speed DC motor with a speed up to 3000 rpm is the basic drive. The sensor used is two piezoelectric accelerometers (CA-YD-106) mounted on the flat surface. The software DASP is used for recording the signals. 77NOVEMBER 2010 The vibration was measured under nine different gear conditions: pattern A-normal condition (no fault), pattern B-single spalling, pattern C-single crack, pattern D-single pitting, pattern E-single tooth broken, pattern F-single gear misalignment, pattern G-compound fault of crack and gear tooth broken, pattern Hcompound fault of wear and gear tooth broken and pattern I-compound fault of wear and spalling, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure15-1.png", + "caption": "Fig. 15 Balanced 1DOF parallelogram derived by fixing link OA of the 2DOF parallel manipulator with idler link", + "texts": [ + " Also for four-bar mechanisms, the different CRCM configurations are applicable just as the substitution of the well-known kinematic relations into the inertia equations of the double and single pendula to obtain the inertia about one of the links. This means that with the equations for the double and single pendula and the kinematic relations the inertia of any four-bar mechanism can be written down easily. It is a special case for which the four-bar linkage becomes a parallelogram. From Fig. 9, and assuming the link between O and A to be fixed with the base, the resulting parallelogram can be balanced, as in Fig. 15, with solely a CRCM. Because the coupler link does not rotate, its center of mass can be located arbitrarily. 5 3DOF Parallel Mechanisms CRCM-balanced 3DOF planar and spatial mechanisms can be synthesized by combining the CRCM-balanced double pendula. Two examples are the planar 3-RRR parallel mechanism of Fig. 16, which has one rotational and two translational DOFs, and the spatial 3-RRR parallel mechanism of Fig. 17, which has two rotational and one translational DOFs. As described in Ref. 16 , the platforms of these mechanisms can be modeled by lumped masses at their joints, maintaining its original mass, its location of the center of mass, and its inertia tensor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003355_1.3653139-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003355_1.3653139-Figure5-1.png", + "caption": "Fig. 5 D i a g r a m s h o w i n g lateral shift of inner a n d outer-ring centers, a l s o notation system < 4 > < T T (17)", + "texts": [ + " 4 Schematic diagram showing out-of-round (exaggerated) inner ring and orientation 4 5 8 / S E P T E M B E R 1 9 6 4 Transactions of the A S M E i I N N E R R I N G GEAR TOOTH LOAD RING ( INTEGRAL GEAR) ING E L E M E N T Downloaded From: http://fluidsengineering.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jfega4/27255/ on 02/15/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use From Fig. 3, it is apparent that the vertical components of loads, both applied and induced, are self-equilibrating owing to symmetry. However, the horizontal loads cause a lateral shift 5i of the initial centers of the inner and outer rings. This is shown in Fig. 5. Considering the lateral shift Si, the relative radial approach of inner and outer-ring raceways at angular positions \u00b14>i is S{ = S, cos , + sCiS + ,CfM + TC/V + Z) pCijP, j P, = Ku> Therefore Uj = 5/ - Tj P, = K(S, - n? FR3 EI o < * < -\u2014 i c o s 0 \u2014 sin $ J pj^Z n 2 ~I \u00a32 = \u2014 \u2014 - \u00a3cos (ir - ) - ^(ir - ) sin (ir - <\u00a3) < 0 < 7T Examination of equations (14) and (17) indicates that maximum clearance occurs at the points of load application; i.e., (j) = \u00b1 i r / 2 and minimum clearance occurs at $ = 0 and ir. Owing to symmetry of loading, only one half of the bearing need be considered in the analysis. In Fig. 5, the rolling-element position nearest the horizontal center line on the left is referenced as i, j = 2. The angular positions , and are then given by (10) Palmgren in [3] states that rolling-element load is a nonlinear function of contact deformation v. This function is 4>i = % (\u00bb - 2 + k) Si 2ir i = - (J - 2 + k) ( 18) (11) in which 6 = 3 /2 for ball bearings and 10/9 for roller bearings. But, contact deformation u is a function of the relative radial approach Sj and the initial radial spacings /', between the rings as follows: in which k = 0 if the first rolling element lies on the horizontal center line; otherwise k = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000532_3-540-36460-9_31-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000532_3-540-36460-9_31-Figure3-1.png", + "caption": "Fig. 3. The retraction mechanism Fig. 4. The accordion effect", + "texts": [ + " 2 still applies during the retraction phase. But in this case, the distal clamp is activated as the bellow retracts. 1l1, the effective retraction stroke, now represents the distance traversed by the proximal end after slipping has occurred. Thus, the inchworm retraction efficiency \u03b7r can be written as: \u03b7r = 1l1 1lt = 1 \u2212 1l0 1lt . (8) Equations (3), (4) and (5) still apply but l0 in (4) now refers to the distance between the proximal end and the nearest proximal mesentery muscle that holds the GI tract in place (Fig. 3). If the device is not a wireless system, much of the generated propulsion force would be used to pull the power supply cable or \u2019tail\u2019 as the device propels itself forward. Together with the weight of the device, this pulling force contributes to the magnitude of Fc. In comparison, Fc in the retraction phase would be much greater than that of the elongation phase. Thus, from (7), the retraction efficiency is derived as \u03b7r = 1 \u2212 l0 1lt \u221a Fc \u03c0\u03b38t . (9) The authors\u2019 past in vivo experimental results (Carrozza et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003355_1.3653139-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003355_1.3653139-Figure2-1.png", + "caption": "Fig. 2 Planet-gear bearing showing gear-tooth loading", + "texts": [ + " In fact, theoretical analysis of a planetary-gear bearing having an out-ofround inner ring, wherein the major axis is oriented parallel to the tangential gear-tooth load, indicates that a condition of optimum clearance and out-of-round exists with respect to bearing fatigue life. An increase in fatigue life of approximately 40 percent is indicated with only a nominal amount of out-of-round. Introduction IN planetary-gear transmissions, and especially for aircraft applications, the planet gears are frequently made integral with the outer rings of rolling bearings. Fig. 1 shows a typical planet gear-bearing assembly. Thus the outer ring of the bearing is subjected to the gear-tooth loads as shown in Fig. 2. Since loads on spur-gear teeth may be resolved into tangential and radial components, the loading on the bearing outer ring may be idealized as shown in Fig. 3. The inner ring of the bearing is loaded only by the rolling-element loads. The inner ring is generally a stiffer structure than the outer ring and the inclusion of inner ring deflections in a fatigue-life analysis would tend to cause smaller rolling-element loads and, hence, longer bearing fatigue life. Thus the omission of the inner ring structure from the analysis causes the results to be conservative regarding bearing fatigue life" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002505_12.918523-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002505_12.918523-Figure2-1.png", + "caption": "Fig. 2 Interception for minimal impact to the base satellite based on the contact force direction", + "texts": [ + " As we have stated early, the overall goal for the control design is to capture the tumbling satellite with minimal impact on the attitude motion of the servicing satellite. The first step for achieving such a goal is to determine a best time for the robot to grasp. It is understandable that if the resultant contact force exerted at the robot tip (resulting from a capture action) passes the mass center of the servicing system, the contact force will not cause any attitude disturbance to the servicing satellite, as shown in Fig. 2. However, the direction of contact force depends on the relative velocity, contacting spots and contact geometry, which make it very difficult to predict in advance. Although such a prediction may not be impossible if we have an accurate contact dynamics model, this will require more extensive research work in the future. For this work, we approximate it by assuming that the contact force is along the direction of the relative velocity between the robot tip and the grasping handle of the target satellite" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001743_3.30282-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001743_3.30282-Figure4-1.png", + "caption": "Fig. 4 Geometric description of modal amplitudes in the missile-fixed plane coordinate system.", + "texts": [ + " Since the projectiles under discussion herein are spinning slowly and only short portions of the trajectory are considered, the to term due to gravity can be dropped. If the motion were D ow nl oa de d by U N IV E R SI T Y O F M IC H IG A N o n Fe br ua ry 1 9, 2 01 5 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /3 .3 02 82 APRIL 1971 BEHAVIOR OF PROJECTILES OVER LONG FLIGHT PATHS 383 tricyclic, a Ks mode with phase <\u00a33 would be needed in addition to the KI and K2 modes. The KI and Kz modal arms lie in the 7-Z plane of a missile-fixed plane coordinate system where the X axis is the missile axis. This representation is shown in Fig. 4. The vectors, Eq. (3), are expressed in a range coordinate system. The transformation from range coordinate system to missilefixed plane coordinate system is accomplished by [1- [1- ( K.V(1 - where = \u00a3* + = \u00a3*\u00bb + *\u00a3.. == (5) (6) (7) (8) (9) Using Eq. (2) with the transformation relations (5-9), an exact but complicated relation between a, us, solar vector, and the modal arms can be derived. This complicated relation can be simplified, ignoring terms greater than third order and the following expression resultsn cosi (0i - 8) cos\u00ab>2 - (10) The phase angles are stated in terms of initial phase angles and frequencies For small angles, then, we can fit the data by Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000076_j.robot.2008.01.002-Figure16-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000076_j.robot.2008.01.002-Figure16-1.png", + "caption": "Fig. 16. Gait planning using crab gait.", + "texts": [ + " Remember that there is a restriction on the crab angle in (3). Also note that the quadruped robot can have optimal ditch crossing ability if it uses both one-phase gaits and two-phase gaits according to the ditch width. In the first, we present gait planning in which the quadruped robot should change the heading direction with the crab gait. In this case, the ditch width d should be greater than Rx/2 so that the quadruped robot cannot cross the ditch either with onephase or two-phase gait (see Fig. 15). Referring to Fig. 16, the front position of the ditch can be denoted by the coordinate xo on the X -axis of the body coordinate system. For preserving the condition that the robot has as few steps as possible, the quadruped would change its heading direction with the maximum crab angle it may have, i.e., with \u03b1max introduced in (3). Since \u03b1max cannot be greater than or equal to 90\u25e6, the quadruped robot should have straight line going along the X - axis until Position 1 in Fig. 16, and should start the fault-tolerant crab gait with the crab angle \u03b1max, following the subsequent route 2 \u2192 3 \u2192 4. In order to have fewest steps, the robot must have two-phase gaits throughout the trajectory. xc1, the X coordinate of Position 1, is determined as follows. First we draw a straight line from the upper left vertex of the ditch that makes the acute angle \u03b1max with the X -axis, and designate as 1\u2032 the position where the center of gravity is on the intersection point x \u2032 c1. Referring to Fig. 16, we can see that xc1 should be l away from x \u2032 c1 for the contour rectangle not to overlap the ditch: xc1 = x \u2032 c1 \u2212 l. From Fig. 16, x \u2032 c1 is expressed with robot parameters and the ditch coordinate as x \u2032 c1 = xo \u2212 yo/ tan \u03b1max. The offset l can be derived as l = Ry + W tan \u03b1max + Rx . Hence given the ditch coordinate, the X coordinate of the center of gravity from which the fault-tolerant crab gait initiates is obtained as below: xc1 = xo \u2212 ( yo + Ry + W tan \u03b1max + Rx ) . (14) The necessary condition for this gait planning is that xc1 should be non-negative, i.e., xo \u2265 (yo + Ry + W )/ tan \u03b1max + Rx from (14). If this condition fails, the quadruped robot will fall into deadlock" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000281_1.2777476-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000281_1.2777476-Figure1-1.png", + "caption": "FIG. 1. Schematic of a microcontact model between a composite asperity and rigid plane.", + "texts": [ + " An accurate general solution for the elasticplastic adhesion for rough surfaces is still missing. Then in this paper, a presented elastic-plastic adhesion model is presented to solve the adhesive interaction of rough surfaces to improve the existing problem, which combines the MD model with the fractal parameters of the surface. To simplify the problem, the system of two opposing rough surfaces is replaced by a rigid flat smooth and a deforming rough countersurface with roughness equivalent to the effective roughness of the two original surfaces and effective Young\u2019s modulus. Figure 1 shows the schematic of a microcontact model of an asperity with a rigid flat. The deformation of the asperity may be elastic, elastoplastic, or plastic according to the pressure condition. In this adhesion model, an asperity is assumed to be either elastic or fully plastic, i.e., elastic\u2013perfectly plastic material behavior is considered at the asperity. According to the GW model, thea Electronic mail: foxpengxmu@xmu.edu.cn and foxpenghit@hit.edu.cn 0021-8979/2007/102 5 /053510/7/$23.00 \u00a9 2007 American Institute of Physics102, 053510-1 [This article is copyrighted as indicated in the article", + "97 On: Wed, 04 Jun 2014 21:14:46 It is well known that the surface topography is fractal, and then the adhesion model of the asperity should be constructed based on the fractal parameters to ensure that the contact theory is scale independent. Yan and Komvopoulos developed a two-dimensional form of WM function to approximate a fractal function in three dimensions. By making this approximation, the following expression for the truncated height of one deformed spherical asperity against a rigid flat as shown in Fig. 1 is developed:15 = 2GD\u22122 ln 1/2 2a 3\u2212D, 14 where D Ds=D+1 is the fractal dimension of the profile and a is the radius of the truncated microcontact area. The radius of the truncated area of the asperity is a = 1 2 2GD\u22122 ln 1/2 1/ 3\u2212D . 15 Yan states that since the radius of curvature R of each asperity is much greater than the height of the asperity, then this relationship between the curvature and the truncated height can be assumed to be a 2=2R . Then the relationship for the radius of curvature as a function of the truncated area and fractal parameters is R = s D\u22121 /2 25\u2212D D\u22121 /2GD\u22122 ln 1/2 , 16 where s = a 2 is the truncated area of the asperity and can be expressed as s = 24\u2212D D\u22123 /2GD\u22122 ln 1/2 2/ 3\u2212D " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003079_msf.723.332-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003079_msf.723.332-Figure7-1.png", + "caption": "Fig. 7 Machined part. Fig. 8 Immersion angle of two strategies", + "texts": [], + "surrounding_texts": [ + "In this paper, the trochoidal milling of nickel-based superalloy is investigated with machining experiments. Machining results show that trochoidal tool path strategy can reduce tool load and tool wear significantly while remain the productivity in milling nickel-based superalloy. Trochoidal tool path strategy can be further developed to machine more complex surface, such as freeform blade surface or blisks." + ] + }, + { + "image_filename": "designv11_3_0001482_optim.2010.5510564-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001482_optim.2010.5510564-Figure1-1.png", + "caption": "Fig. 1 Dc current decay standstill test wiring diagram for ALA rotor 4 pole RSM", + "texts": [ + " The paper is organized as follows: in Section II \u2013 the ALA rotor RSM parameters are determined, in Section III \u2013 the RSM space-vector model is ilustrated, in Section IV the active flux concept is introduced, in Section V the I-f control method is presented, in Section VI the proposed motion sensorless control of the RSM is described, transition strategies are detailed in Section VII, experiments are shown in Section VIII based on which conclusion emerge in Section IX. II. THE ALA ROTOR RSM PARAMETERS The ALA rotor RSM is credited with high saliency (high performance [2]), and thus, despite of its manufacturing apparent difficulties, is investigated here. The RSM electrical parameters were determined after performing dc current decay tests with the rotor aligned along d axis, and, respectively, along q axis. 337978-1-4244-7020-4/10/$26.00 '2010 IEEE As can be seen from Fig.1 the current decay occurs in phases B and C, which are actually connected in series, while in phase A different values of constant current are injected. The d-q fluxes were calculated using Equation (1): , (2 ) / 2d q DR idt V dt= \u22c5 \u22c5 +\u222b \u222b\u03bb (1) Where: R \u2013 phase resistance; i-decay current, VD-diode voltage drop. The \u03bbd\u2013Id and \u03bbq\u2013Iq dependencies are illustrated in Fig. 2 a and b. The cross coupling saturation, visible mainly in axis d, will not influence notably the proposed control system as, for the active flux estimation, we use the machine voltage model (above 5 Hz) and Lq which is rather constant for an ALA rotor, Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001356_s00205-009-0215-z-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001356_s00205-009-0215-z-Figure7-1.png", + "caption": "Fig. 7. Non-convex pinned sets", + "texts": [ + " The next example shows the existence of polyrectangular sets without convexity properties that are fixed points for our motion (namely, they are pinned). This highlights a difference with the standard crystalline motion where, as for the case of isotropic curvature flow, initial connected sets become convex in finite time (and then shrink to a point). Example 1. [Polyrectangular (non-convex) pinned sets] Consider the initial set E = ([\u2212R1, R1] \u00d7 [\u2212R2, R2]) \u222a ([\u2212R2, R2] \u00d7 [\u2212R1, R1]), where \u03b1 < R1 < R2 (the set on the left in Fig. 7). Then the sides Si of E either have curvature of sign 0, or length larger than 2\u03b1 so that L\u0307i = 0 in the previous theorem, and E(t) = E for all times. Another example is the set on the right in Fig. 7. Note that this set is not even geodesically convex with respect to the distance related to the l1-norm (that is, not all pairs of points are connected by a minimal path with respect to that distance), while the first one is. Example 2 (pinning after an initial motion). Consider as initial set a square of side length larger than 2\u03b1 from which a small square has been removed (see Fig. 8). Then the larger boundary stays pinned while the inner square shrinks to a point after a finite time. After that time the motion is constant (equal to the larger square)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003189_j.cnsns.2011.04.012-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003189_j.cnsns.2011.04.012-Figure11-1.png", + "caption": "Fig. 11. vertical", + "texts": [], + "surrounding_texts": [ + "Due to the high nonlinearity of gas film forces, the system behavior is studied numerically using the finite element method. The flow domain of each lobe of four-lobe bearing is rectangular, therefore, four-noded rectangular isoparametric element is employed to mesh flow field. In this study, the fourth order Runge\u2013Kutta method is used for the integration of Eq. (25). By many trails, time interval of Dt = p/300 is found to be the optimal considering accuracy of the results and computation time. The Gauss\u2013Siedel iteration method is employed for the solution of Eq. (11) to obtain f _Wg column during the time integral scheme. The convergence criterion is applied on every nodal value _Wi as j _Wi;j\u00fe1 _Wi;jj 6 _Wtol i \u00bc 1;2; . . . nf \u00f026\u00de where \u2018j\u2019 is the iteration number. This condition states that no nodal values of _W in the solution field should change by an amount greater than _Wtol as a result of one Gauss\u2013Siedel iteration. In the present work _Wtol is taken to be 10 6. The time series data of the first 600000 time steps are excluded from dynamic behavior investigation to ensure that the data used represent steady state conditions. The resulting data include the orbital paths of the rotor center. These data are then used to generate the power spectra, Poincare maps and bifurcation diagrams. Fast Fourier transformation is used to obtained power spectra of the rotor center in horizontal and vertical directions. To generate Poincare map, a Poincare section that is transverse to the flow of a given dynamic system is considered. A point on this section is a return point of the time series at the constant time interval of T, that T is the driven period of Trajectory of the rotor center at mr=18.07 kg (a); phase portraits of rotor center (b) and power spectra of rotor displacement in horizontal (c) and (d) directions for k = 1.5. the exciting force in non-autonomous systems. The projection of a Poincare section on the x y plane is related to the Poincare map of the dynamic system. Bifurcation diagram is a useful means to observe nonlinear dynamic behavior of a system. To draw bifurcation diagram, an obtained point on the Poincare map is used with varying rotor mass value by a constant step." + ] + }, + { + "image_filename": "designv11_3_0002070_1.1658083-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002070_1.1658083-Figure5-1.png", + "caption": "FIG. 5. The force in nanonewtons exerted by a defect on a wall vs time in nanoseconds for the material parameters listed in Table I. Step of applied field.", + "texts": [ + " The pressure Puis now unsaturated and depends on z. It is given by (+'0/2 Pu=Nk L (z'-z)dz'= -NkzQz= -Nfoz. -zo/2 (11) The pressure exerted by the defects on the wall is then as shown by the full line in Fig. 4. Wall motion within the limits -zo/2':::;z':::;+zo/2 is reversible and has no hysteresis. If the wall is moved past zo/2, say to point f3z= 2M.H+P(z) , (12) where f3 is a viscous-damping constant, H is the applied field, M. is the saturation magnetization, and P(z) is given in Fig. 4. It may be noted that Fig. 4 is identical with Fig. 5 of Ref. 8. As a consequence, all of the results derived in that paper follow from the present model. In particular, it predicts that (1) There is a reversible spike of wall velocity at the leading and trailing edges when a rectangular pulse of field is applied to a wall. (2) The steady-state velocity in a constant field is proportional to H - He, where H. is the threshold field for irreversible wall motion, as shown in Fig. 2(b). (3) The small-signal hysteresis loop is Rayleigh-like if the distribution of defect strengths is essentially constant for small values of zoo The present model is more satisfactory than that of Ref", + " Also Ht 2 1t - = J(A)g'(t-A)dA 2 0 (42) from (40). Finally transformation of (24) into the time domain gives (3tmn(t) = [amn(A) exp[-smn(t-A)]dA. (43) o First let) was evaluated from (41). ThenJ(t) was found by numerical deconvolution of (42). Using this, the amn(t) were determined from (22). This allowed the Fourier coefficients tmn (t) to be found using (43). Finally Wet) and z(t) were evaluated using (37) and (31) . The force J exerted on the wall by a defect is shown as a function of time in Fig. 5. Both the result of the exact calculation and the prediction of the spring model are shown. The maximum deviation between the two is about 7% of the final value. The average wall position z as a function of time is shown in Fig. 6. The wall position calculated from the spring model is also shown. The maximum deviation is about 3.5% of the final value. The energy W associated with a single defect [This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000844_tmag.2007.916121-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000844_tmag.2007.916121-Figure7-1.png", + "caption": "Fig. 7. Real part of the deformation at 3333 Hz.", + "texts": [], + "surrounding_texts": [ + "The better correlation of the body sound index for 3333 and 6000 Hz can be understood looking at the mode shapes at both frequencies in Figs. 7 and 8. It is clearly seen that the deformation correlates better in terms of shape and magnitude in the 3333-Hz case than in the 6000-Hz case. A closer look at the 3-D solution shows that the 3333-Hz case is an almost pure 2-D mode shape, where the deformation at 6000 Hz varies strongly along the axial direction of the machine. A previous study [4], however, showed that 6000 Hz was one of the most dominant frequency regarding the acoustic noise radiated by this SRM. Therefore, it can be concluded that the particular caution has to be taken when analyzing electrical machines by means of 2-D structure-dynamic simulation, especially if the machine is face-mounted." + ] + }, + { + "image_filename": "designv11_3_0001764_s11036-009-0152-y-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001764_s11036-009-0152-y-Figure7-1.png", + "caption": "Fig. 7 A simple coordination graph, the action sets A1, A2, A3, the atomic and pairwise payoffs f1, f2, f3, f12, f23 and the global utility function u", + "texts": [ + " Generally the goal is to find the optimal action vector a\u2217 = (a\u2217 1, . . . , a\u2217 n) that maximizes a global utility function u(a). The utility has a structure captured by a coordination graph CG = (V, E). On every node of CG we define a function fi(ai) called atomic payoff. An atomic payoff describes how well each action serves the goal of the agent corresponding to that node. On the edges eij of CG we define pairwise payoff functions fij(ai, a j) that indicate how good for the team are pairs of actions of interacting agents (see Fig. 7 for an example). The global utility is assumed to depend only on the unary and pairwise payoff functions as follows: u(a) = \u2211 i fi(ai) + \u2211 eij\u2208E(CG) fij(ai, a j) Max-plus is a distributed message passing algorithm that attempts to compute an optimal action vector using only local computations and communication for every agent. While the algorithm is running, each agent i chooses a neighboring agent j on CG, collects and adds all incoming messages from other agents in its neighborhood, and sends a new message to j that is computed by the following formula: mij(a j) = max ai \u23a7 \u23a8 \u23a9 fi(ai) + fij(ai, a j) + \u2211 k\u2208N(i),k = j mki \u23ab \u23ac \u23ad (7) At any time during the execution, the agents can compute a marginal function gi(ai) = fi(ai) + \u2211 k\u2208N(i) mki" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001848_1755952.1755975-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001848_1755952.1755975-Figure2-1.png", + "caption": "Figure 2: Biped dimensions, point-mass locations and measuring conventions.", + "texts": [], + "surrounding_texts": [ + "In this first application we emphasize two ideas. First, the following two-domain hybrid system is an example of a system that can have hybrid Poincare\u0301 maps with rank strictly less than the upper bound derived in Theorem 4. Secondly, computations will show that the rank of both Poincare\u0301 maps is equal to 0. This means that the hybrid limit cycle is not sensitive to perturbations away from it; such systems are sometimes called superstable, or are said to display instantaneous convergence to a limit cycle. We define the first domain, D1, of this two-domain hybrid system to be the upper-right quadrant of R 2. The vector field in this domain is f1(x, y) = (\u2212y + x(1\u2212 x2 \u2212 y2), x+ y(1\u2212 x2 \u2212 y2) )T . The flow resets to the next domain when it reaches the positive y-axis, which we define to be the guard Ge1 . The reset map Re1 projects the y-axis into R 3 such that Re1(0, y) = (0, y, 0)T clearly has rank 1. The flow in the second domain, D2, is the linear system f2(x, y, z) T = (\u2212x,\u2212z, y)T , and is permitted to flow from the y-axis in R 3 until it reaches the xz-plane, which defines Ge2 . All points on the xz-plane are mapped back to the first domain by the reset map Re2(x, 0, z) = (x+ 1, 0)T , which also has rank 1. Considering Theorem 4 and Lemma 3 together yields 0 \u2264 rank(P1) \u2264 1, 0 \u2264 rank(P2) \u2264 1, where P1 is the Poincare\u0301 map defined from Ge1 and P2 likewise from Ge2 . We find that rank(P1) = rank(P2) = 0, through numeric computation. Therefore, equality with the upper bound of Theorem 4 is not obtained. In addition, we may interpret this to mean that all trajectories emanating from Ge1 or Ge2 will converge to the limit cycle after at most one iteration \u2014 see Figure 1." + ] + }, + { + "image_filename": "designv11_3_0000183_tmag.2008.2002996-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000183_tmag.2008.2002996-Figure2-1.png", + "caption": "Fig. 2. Reluctance network models of the stator and the massive rotor or flux barrier rotor, and interconnection principle of the stator and rotor networks.", + "texts": [ + " The flux density is also assumed to be radial near the stator inner surface. Thus, the airgap length function is separable into two independent parts and , respectively, linked to the stator and the rotor, measured relatively to a borderline located in the middle of the airgap. and are periodic functions with respective periods of one slot pitch and one pole pitch. has a fixed spatial origin, but the origin of varies with the position of the rotor. For a given position of the rotor, the total airgap length function is then computed by (1) 0018-9464/$25.00 \u00a9 2008 IEEE Fig. 2 shows the reluctance networks of the stator and of both rotors. Nonlinear reluctances are colored in black. The thick nodes show the interconnection nodes. The stator reluctance network includes the MMF sources at each tooth. Each interconnection node is associated to an angular zone with a span of a tooth pitch and centered in the middle of the tooth tip. The massive rotor is divided into four reluctances in order to allow the flux issues to flow out of the rotor. We found that this discretization leads to results in good accordance with the finite-element ones for any value of the load angle", + " The remaining area at each rotor side is divided into three zones with equal spans. These zones are associated with the three middle nodes. For the flux barrier rotor, we also found that a longitudinal subdivision of each ferromagnetic segment into four reluctances in series and a subdivision of each flux barrier into three reluctances in parallel lead to a good agreement with the finite-element results. is the nonlinear total direct reluctance of the ferromagnetic segment number , and is the linear \u201ctotal\u201d quadrature reluctance of the flux barrier number . As shown in Fig. 2, the flux barrier reluctance is distributed on the three parallel paths so that the reluctance of the middle path is twice lower than the two others. In the flux barrier rotor, only the external nodes may be interconnected to the stator nodes. On each rotor pole side, the angular zone associated with an interconnection node is centered in the middle of the corresponding ferromagnetic segment tip. It is delimited by the middles of the adjacent flux barrier tips. Similarly to the massive rotor, each area facing a lateral side of the rotor is divided into three angular zones with equal spans of . These zones are associated with the three middle interconnection nodes at each rotor lateral side. For both rotors, only the direct flux path can be saturated, because the quadrature flux is generally too low to cause saturation. For any rotor position, the reluctance network modeling the airgap interconnects the rotor and the stator reluctance networks into one global network. Fig. 2 explains the principle of building the interconnection network. The interconnection nodes (thick nodes) and their associated angular zones are represented. At each time step, the rotor position changes and the airgap network has to be recomputed. For each rotor interconnection node, a subroutine determines if there is a connection or not to one or more stator interconnection nodes. The stator and rotor nodes of respective numbers and are connected if their angular zones overlap. The value of the permeance connecting them is calculated by integrating the inverse of the total airgap length function over the overlapping angular zone (2) where is the active length of the motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003215_j.cad.2011.03.004-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003215_j.cad.2011.03.004-Figure10-1.png", + "caption": "Fig. 10. CAD drawing of the designed test part. The numbers indicate the five features of the part.", + "texts": [ + " To provide proof of concept for the robustness of the proposed beam compensation tool, it was integrated in an in-house software tool for complete jobpreparation for additivemanufacturing. A test part was designed, processed and built on an in-house Selective Laser Melting apparatus [42,6]. The apparatus has an Yb\u2212 Yag fiber laser. The atmosphere in the building chamber is made inert with argon gas. The material used is the titanium alloy Ti6Al4V, and the process parameters are 45 W laser power and a scan speed of 180 mm/s. The layer thickness was 0.03 mm. The test part, shown in Fig. 10, consisted roughly of three vertical limbs connected at the bottom by a thin pedestal. The following features were present in the designed part: 1. A leftward-oriented horizontal cylindrical beam, supported by a triangle of which the thickness varied from 0.150\u20130.300 mm from bottom to top; 2. A forward- and a backward-oriented horizontal cylindrical beam, supported by a triangle of which the thickness varied from 0.100 to 0.200 mm from bottom to top; 3. Themiddle limb consisted of concave beams ofwhich thewidth was 300 mm at top and bottom of the beam and 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002495_s13198-013-0173-6-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002495_s13198-013-0173-6-Figure3-1.png", + "caption": "Fig. 3 The two tapings connected to the stator coils of the phase W1\u2013 W2", + "texts": [ + " In the case of broken rotor bars, the PVMP\u2013H becomes: PVMP H \u00bc A0 \u00fe Am TH cos m 1\u00f0 \u00dexstf g \u00fe Am RSH cos m 1\u00f0 \u00dexs kNrxr\u00bd tf g \u00fe Am RBFH cos m 1\u00f0 \u00de 2ks\u00bd xs tf g \u00f024\u00de In the case of motor with eccentricity fault, the PVMP\u2013H becomes: PVMP H \u00bc A0 \u00fe Am TH cos m 1\u00f0 \u00dexstf g \u00fe Am RSH cos m 1\u00f0 \u00dexs kNrxr\u00bd tf g \u00fe Am EFH cos m 1\u00f0 \u00dexs kxr\u00bd tf g \u00f025\u00de For a motor with inter-turn short-circuit, the PVMP\u2013H can be expressed as: PVMP H \u00bc A0 \u00fe Am TH SC : cos m 1\u00f0 \u00dexstf g \u00fe Am RSH SC : cos m 1\u00f0 \u00dexs kNrxr\u00bd tf g \u00f026\u00de This theoretical study shows clearly that the signal of PVMP\u2013H contains potential information of motor state and its spectral analysis can help us to make a good diagnosis. The test bench used in the experimental investigation is available in the LGEB at university of Biskra-Algeria (Fig. 2). The motor exploited to study the occurrence of inter-turn short-circuit faults is a three-phase 50-Hz, Y connection, 4-pole, 3 kW Leroy Somer squirrel-cage induction machine. The stator winding was modified by the addition of two tapings connected to the stator coils of the phase W1\u2013W2 (Fig. 3). The other ends of these external wires are connected to the motor terminal box that allows introducing inter-turn short circuits with different number of turns. The motor was tested at healthy state, with 4 shorted turns (2 %) then with 10 shorted turns (5 %) under different loads. In all cases, 10 s of the three-phase currents were sampled at 10 kHz using a dSpace 1104 card. After acquisition, the six steps of the proposed approach were applied to calculate the PVMP\u2013H. Then, all data were processed using the MATLAB software package to compute the fast Fourier transform (FFT)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003373_acc.2012.6315362-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003373_acc.2012.6315362-Figure2-1.png", + "caption": "Fig. 2. Schematic representation of the Qball-X4", + "texts": [ + " A commonly employed quadrotor UAV model [11] is: mx\u0308 = uz (cos\u03c6 sin\u03b8 cos\u03c8 + sin\u03c6 sin\u03c8) ; J1\u03b8\u0308 = u\u03b8 my\u0308 = uz (cos\u03c6 sin\u03b8 sin\u03c8\u2212 sin\u03c6 cos\u03c8) ; J2\u03c6\u0308 = u\u03c6 (1) mz\u0308 = uz (cos\u03c6 cos\u03b8)\u2212mg; J3\u03c8\u0308 = u\u03c8 where x, y and z are the coordinates of the quadrotor UAV center of mass in the earth-frame. \u03b8 , \u03c6 and \u03c8 are the pitch, roll and yaw Euler angles respectively. m is the mass and Ji (i = 1,2,3) are the moments of inertia along y, x and z directions respectively. uz is the total force generated by the four propellers and applied to the quadrotor UAV in the zdirection (body-fixed frame). u\u03b8 , u\u03c6 and u\u03c8 are respectively the applied torques in \u03b8 , \u03c6 and \u03c8 directions (see Figure 2). The relation between uz, u\u03b8 , u\u03c6 , u\u03c8 and the PWM inputs to the four motors is given as: uz = K(u1 +u2 +u3 +u4) u\u03b8 = KL(u1\u2212u2) u\u03c6 = KL(u3\u2212u4) u\u03c8 = KK\u03c8(u1 +u2\u2212u3\u2212u4) (2) where K and K\u03c8 are constants and L is the distance from the center of mass to each motor. A simplified linear model will be employed for trajectory planning. This model can be obtained by assuming hovering conditions (uz \u2248 mg in the x and y directions), no yawing (\u03c8 = 0) and small roll and pitch angles: x\u0308 = \u03b8g; J1\u03b8\u0308 = u\u03b8 y\u0308 =\u2212\u03c6g; J2\u03c6\u0308 = u\u03c6 (3) mz\u0308 = uz\u2212mg; J3\u03c8\u0308 = u\u03c8 One objective of the trajectory planning is to reduce energy consumption by driving a system as fast as possible from one position to another" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002632_j.mechmachtheory.2013.02.010-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002632_j.mechmachtheory.2013.02.010-Figure5-1.png", + "caption": "Fig. 5. Gear pair model with two pinions.", + "texts": [ + " The load sharing index (LSI) can be defined as the ratio of the minimal and maximal Floquet multipliers LSI \u00bc min \u03bbij j=max \u03bbij j; i \u00bc 1; \u22ef;n: \u00f024\u00de The LSI describes the dynamic behavior of transmission errors of every mesh gear pair. The range of LSI is from 0 to 1. The load sharing behavior becomes worse when LSI gets smaller. In this case, the mesh force of some gear pair is bigger than that of others, which is dangerous when related to damage and failure of structures. On the other hand, when the LSI increases, it means that the load sharing behavior becomes better. The load on every mesh gear pair is then similar. Here, an example gear system where one gear is driven by two pinions as shown in Fig. 5, is employed to investigate the correlation of the load sharing behavior and various mechanical parameters. In the model, the elastic deformation of the mounted constraints and the relative motions on the meshing surfaces are considered. The dynamic equations of motion are mp1\u20aczp1 \u00bc \u2212cbp1 _zp1\u2212kbp1zp1\u2212cm1 Rp1 _\u03b8p1\u2212Rg _\u03b8g \u00fe _zp1\u2212 _zg \u2212km1 t\u00f0 \u00de Rp1\u03b8p1\u2212Rg\u03b8g \u00fe zp1\u2212zg1 ; mp2\u20aczp2 \u00bc \u2212cbp2 _zp2\u2212kbp2zp2\u2212cm2 Rp2 _\u03b8p2\u2212Rg _\u03b8g sin \u03b32 \u00fe _zp2\u2212 _zg \u2212km2 t\u00f0 \u00de Rp2\u03b8p2\u2212Rg\u03b8g sin\u03b32 \u00fe zp2\u2212zg ; mg\u20aczg \u00bc \u2212cbg _zg\u2212kbgzg \u00fe cm1 Rp1 _\u03b8p1\u2212Rg _\u03b8g \u00fe _zp1\u2212 _zg \u00fe km1 t\u00f0 \u00de Rp1\u03b8p1\u2212Rg\u03b8g \u00fe zp1\u2212zg \u00fe cm2 Rp2 _\u03b8p2\u2212Rg _\u03b8g sin \u03b32 \u00fe _zp2\u2212 _zg \u00fe km2 t\u00f0 \u00de Rp2\u03b8p2\u2212Rg\u03b8g sin\u03b32 \u00fe zp2\u2212zg ; Ip1\u20ac\u03b8p1 \u00bc \u2212cm1Rp1 Rp1 _\u03b8p1\u2212Rg _\u03b8g \u00fe _zp1\u2212 _zg \u2212km1 t\u00f0 \u00deRp1 Rp1\u03b8p1\u2212Rg\u03b8g \u00fe zp1\u2212zg \u00fe Tp1; Ip2\u20ac\u03b8p2 \u00bc \u2212cm2Rp2 Rp2 _\u03b8p2\u2212Rg _\u03b8g \u00fe _zp2\u2212 _zg = sin \u03b32 \u2212km2 t\u00f0 \u00deRp2 Rp2\u03b8p2\u2212Rg\u03b8g \u00fe zp2\u2212zg = sin \u03b32 \u00fe Tp2; Ig\u20ac\u03b8g \u00bc cm1Rg Rp1 _\u03b8p1\u2212Rg _\u03b8g \u00fe _zp1\u2212 _zg \u00fe km1 t\u00f0 \u00deRg Rp1\u03b8p1\u2212Rg\u03b8g \u00fe zp1\u2212zg \u00fe cm2Rg Rp2 _\u03b8p2\u2212Rg _\u03b8g \u00fe _zp2\u2212 _zg = sin \u03b32 \u00fe km2 t\u00f0 \u00deRg Rp2\u03b8p2\u2212Rg\u03b8g \u00fe zp2\u2212zg = sin \u03b32 \u2212Tg : \u00f025\u00de Substituting q1 = Rp1\u03b8p1 \u2212 Rg\u03b8g + zp1 \u2212 zg, q2 = Rp2\u03b8p2 \u2212 Rg\u03b8g + (zp2 \u2212 zg)/sin \u03b32 into Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002362_tmag.2012.2199094-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002362_tmag.2012.2199094-Figure6-1.png", + "caption": "Fig. 6. Illustration of the 2D models used to calculate the force and torque separately. (a) 2D model definition in the axisymmetric coordinate system to calculate the magnetic fields for translation. (b) 2D model definition in the polar coordinate system to calculate the magnetic fields for rotation.", + "texts": [ + " However, as the windings are orthogonal and oriented in the and the direction, 2D models in the axisymmetric and polar coordinate system can be exploited to approximate the magnetic fields. Therefore, two semi-analytical models are derived based on Fourier analysis as presented in [10] to calculate the translational and rotational performance. A. 2D Model in Axisymmetric Coordinate System A 2D model in the axisymmetric coordinate system is derived to approximate the translational performance of the actuator. The model, as illustrated in Fig. 6(a), is a simplification of the structure as shown in Fig. 5. Due to the selected coordinate system, the model is invariant in the circumferential direction. As such, the slots in the axial direction are not taken into account, hence, the stator back-iron is modeled as a soft-magnetic cylinder with infinite permeability. Three cylindrical concentric regions are defined as shown in Fig. 6: I The non-magnetic shaft of the mover. II The PM-array of the mover. III The airgap and slotless winding for translation. To calculate the magnetic fields due to the PM-array and the coils, the magnetic vector potential is exploited. The formulated model is typically used to calculate magnetic fields in TPMAs as described in [8], [10]. To create a 2D approximation of the PM-array in the axisymmetric coordinate system, the mean value of the components of the magnetization vector along the circumference is calculated", + " Conversely, the axial component of the magnetic flux density is approximately 4% lower in the 2D model. This component can be considered as flux leakage and is more significant in the 3D structure due to presence of the slots. If the width of the slots is increased, the difference in the resulting magnetic field in the two models will increase and vice versa. As such, the 2D model can be used to approximate the performance in the translational direction while the 3D model can be used to calculate the performance reduction due to the presence of the slots. B. 2D Model in Polar Coordinate System Fig. 6(b) illustrates the two dimensional model in the polar coordinate system to calculate the performance of the rotational part of the actuator. Region I, II, III are the same as for themodel in the axisymmetric model. Two regions are added to represent the slot-openings (region IV) and the slots for the winding for rotation (region V). As the stator back-iron is invariant in the axial direction, the stator structure is not simplified in this 2D representation. To approximate the PM-array in the polar coordinate system, the mean value of the two components of the magnetization vector in the axial direction over is calculated as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003583_iros.2011.6094491-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003583_iros.2011.6094491-Figure6-1.png", + "caption": "Fig. 6. Two-spoke rimless wheel model of a biped. (a) Single contact phase: I is spoke length, P is the projected CoP about which the biped is rotating, e is the angle made by the support spoke with the vertical and a, called the leg angle, is the half of the angle between the spokes. (b) Double contact phase: PI and P2 are the two contact points. ,),1 and ,),2 are the slope angles. 0 is the point of intersection of the slopes, and f3 is the angle made by P1P2 with the horizontal.", + "texts": [ + " The central idea behind the controller is as follows: when the high level controller decides to execute a step due to a large push, the reactive stepping controller computes the stepping location where the rimless wheel model would come to a complete stop with its CoM directly above the step location as shown in Fig. 3(c). We call this stepping point the generalized foot placement estimator (GFPE). This is an extension of the FPE [7]. Note that we first compute the GFPE and the corresponding configuration of the rimless wheel will be created with its leg angle a defined as the half of the angle between the two spokes (See Fig. 6), which is computed according to the GFPE. Since the rimless wheel model is defined in 2D and the robot exists in 3D, we determine a plane on which the simplified rimless model resides. We assume that the robot will move in the same direction as the velocity of the robot CoM, Te, just after a push. Consequently, the GFPE will be located on the line of Te projection on the ground. As shown in Fig. 4, the 2D plane for the rimless wheel is defined by two vectors: Te and the vertical through the CoM", + " The rotational inertia of the model is equal to the rotational inertia of the robot about P. More details of the model can be found in [16]. For discussion of the model we assume a flat ground, however our stepping strategy is valid for general non-level ground. We review the dynamic equations for each of the four phases in the scenario of Fig. 5. The equation of motion for a rimless wheel is that of an inverted pendulum when it is in the single contact phase (first and third phases), shown in Fig. 6(a). Ipe = mgl sin e, (3) where e is the angle made by the support spoke and the vertical as shown in Fig. 6(a), Ip is the rotational inertia of the rimless wheel about P, m is mass, and g is the acceleration due to gravity. Until stepping of the swing leg, the total mechanical energy of the rimless wheel E is conserved. At time C just before the touchdown collision, the con servation of E leads to the following relationship: 8(C) = \u00b1 2 -(E - mgl cos a) , Ip (4) since e(c) = a. The sign of 8(C) is decided by the initial condition. For example, in the convention of Fig. 6(a), the sign is negative. During the collision in the second phase, we assume that the angular momentum of the model, kp, is conserved around the collision point. In other words, (5) where t+ is the time just after the collision. The collision results in an instantaneous loss of angular velocity given by: 8(t+) = 8(C) cos(2a) . (6) Based on this dynamic model, we next determine the GFPE for reactive stepping on non-level ground. On non-level ground with discrete slope change as shown in Fig. 6(b), the robot must decide if it should take a step on the first slope or the second slope. Introducing the constants d1 = P10 and d2 = OP2, where 0 is the point of intersection of the slopes, PI is the anchor point and P 2 is the touch-down location of the swing spoke on the second slope, the leg angle ao for stepping on 0 is determined as follows: 21 sin ao = d1. (7) The robot should take a step on the second slope if any of the two following conditions are satisfied: 1) After the step at time t+, when the pivoting spoke just detached from the first slope, e(t+) = a - /1 is negative", + " Another geometric limit is the case that even 90-degree a, presumably the maximum leg angle, cannot stop toppling as shown in Fig. 7(b). Note that in most cases we would not have these limitations since the two cases implies an extreme situation such as the slope is too steep or the robot has already fallen too much. 2 For the sake of convenience, from now on we exclude the two cases in Fig. 7. If we can adjust the length of the swing spoke, which is possible for a humanoid, this limitation may be overcomed in some cases. This will be considered in the future. When the robot steps on the second slope as shown in Fig. 6(b), the angle 8 between the support spokes and the vertical line are: 8(C)=a-(3 8(t+) = a + (3. (10) The distance from 0 to the stepping point, P2 in Fig. 6(b) can be computed as: d2 = -d1 cosh1 + /2) + )4[2 sin2 a - di sin 2 h1 + /2), (11) and (3 is given as: (3 d1 cos /1 + d2 cos /2 cos = \ufffd--\ufffd\ufffd--\ufffd--\ufffd 2l sin a . (3 -d1 sin /1 + d2 sin /2 sIn = ------::-':-.,-----=--\ufffd:::. 2l sin a . Note that d2 and (3 are functions of a only. (12) After the step, we expect the rimless wheel to completely stop at the top as in the right of Fig. Sea). By virtue of conservation ?f energy between the third and fourth phases of Fig. Sea), 8(t+) can be obtained as follows: 2mgl -1- [1 -cos(a + (3)J" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001153_j.jbiomech.2008.01.021-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001153_j.jbiomech.2008.01.021-Figure1-1.png", + "caption": "Fig. 1. Computational simulation models. (a) Experiments 1\u20132: coordinates, weights and load, and resultant moment at joints, represented by semi-filled arrows. Model refers to upper and lower arm, from shoulder (angle q2, control moment c2), via elbow \u00f0q1; c1\u00de to wrist, where load from hand is attached: gravity g3, and carried load p1. L1 is length of lower arm, L2 of upper; g1 and g2 are corresponding gravity loads. (b) Experiment 4: four segment sagittal model of high jumping, with segment lengths, angles and mass moments of inertia marked. Lengths L01, L02, L03, L04 define distances to mass center of segment from its distal end.", + "texts": [ + " Choosing any of the other criteria to express a certain aspect of movement, might demand inclusion of at least a small acc, to numerically resolve the redundancy. In the example below, the a coefficients were set non-zero, one at the time, and the case denoted by the non-zero coefficient, e.g., an acc solution indicates that acc40, with all the other coefficients zero; it is noted if the non-zero a is not unity. ARTICLE IN PRESS M. Kaphle, A. Eriksson / Journal of Biomechanics 41 (2008) 1213\u20131221 1215 Arm movement in the sagittal plane was studied. The model consisted of two linked rigid segments with joint angles q1 and q2 as degrees of freedom, Fig. 1a; the hand was not considered. Anatomical data were: L1 \u00bc 0:32m, L2 \u00bc 0:25m, g1 \u00bc 11N, g2 \u00bc 20N, g3 \u00bc 5N, based on anthropometrical data of a 50th percentile male (Winter, 2005). The weights g1 and g2 acted at the centers of the corresponding segments, while g3 and external load p1 \u00bc 10N acted at the wrist. The dynamic equilibrium equations were obtained using the Euler\u2013Lagrange equations. Two mo- ments c1 and c2 acted at the elbow and shoulder joints, respectively. Moment arm data for the muscles were obtained from an upper extremity model (Holzbaur et al", + " A similar setting as in experiment 1 was used, but without gravity and external forces. No restrictions were introduced for moments. Elbow hyperextension was restricted by 0pq1\u00f0t\u00de. Simulations were performed for two different movements in the horizontal plane, according to Uno et al. (1989). Three different criteria were used. A four link sagittal model was used to simulate high jumping with equations and relevant data from Pandy et al. (1990). The model used four segment angles as coordinates \u00f0q1; q2; q3; q4\u00de, Fig. 1b. Data were chosen from the reference as: m1 \u00bc 2:2 kg, m2 \u00bc 7:5 kg, m3 \u00bc 15:15 kg, m4 \u00bc 51:22 kg, L1 \u00bc 0:175m, L2 \u00bc 0:435m, L3 \u00bc 0:400m, L01 \u00bc 0:095m, L02 \u00bc 0:274m, L03 \u00bc 0:251m, L04 \u00bc 0:343m, I1 \u00bc 0:008 kgm2, I2 \u00bc 0:065 kgm2, I3 \u00bc 0:126 kgm2, I4 \u00bc 6:814 kgm2. The phase from heel lift-off to body lift-off was analyzed. The high jump started from a crouching position, q1\u00f00\u00de \u00bc 34 , q2\u00f00\u00de \u00bc 120 , q3\u00f00\u00de \u00bc 30 , q4\u00f00\u00de \u00bc 120 , and all the segments except the feet were vertical when the lift-off took place: q1\u00f0T\u00de \u00bc 60 , q2\u00f0T\u00de \u00bc q3\u00f0T\u00de \u00bc q4\u00f0T\u00de \u00bc 90 with zero initial velocities" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003348_tie.2012.2205353-Figure19-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003348_tie.2012.2205353-Figure19-1.png", + "caption": "Fig. 19. Equivalent circuit of the short-circuit path.", + "texts": [ + " (A3) Reorganizing terms in (A3) yields m N ( N d\u03c6A dt \u2212RsiA ) \u2212m N Rsif\u2212 m2 N2 Lls dif dt =Rf if . (A4) To get rid of the flux \u03d5A and the phase current iA, substitute (A2) into (A4) m N ( vab + m N Rsif + m2 N2 Lls dif dt ) \u2212m N Rsif \u2212 m2 N2 Lls dif dt = Rf if . (A5) Combining terms in (A5) yields m N vab \u2212 ( 1\u2212 m N ) m N Rsif \u2212 ( 1\u2212 m N ) m2 N2 Lls dif dt = Rf if . (A6) Equation (A6) establishes a direct relationship between vab and if . Equation (A6) can be described by an equivalent circuit, as shown in Fig. 19. The equivalent circuit in Fig. 19 can be used to simulate the fault current if regardless of the number of shorted turns. In the case that m N , the term 1\u2212m/N in Fig. 19 can be simply replaced by one. The resistance term can also be absorbed into the fault resistor Rf . Consequently, the equivalent circuit in Fig. 19 can be further simplified to the equivalent circuit in Fig. 3 when m N . In the extreme case that m = N , m/N is equal to one, and the term 1\u2212m/N is zero. Fig. 19 then shows that the voltage source vab is directly connected to the resistor Rf . This describes exactly what happens when a resistor is connected across the entire phase. REFERENCES [1] P. O\u2019Donnell, \u201cReport of large motor reliability survey of industrial and commercial installations, Part I,\u201d IEEE Trans. Ind. Appl., vol. IA-21, pp. 853\u2013864, Jul. 1985. [2] O. V. Thorsen and M. Dalva, \u201cA survey of faults on induction motors in offshore oil industry, petrochemical industry, gas terminals, and oil refineries,\u201d IEEE Trans" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001785_j.mechatronics.2009.01.003-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001785_j.mechatronics.2009.01.003-Figure1-1.png", + "caption": "Fig. 1. The kinematic model.", + "texts": [ + " (14) can be rewritten as JT s\u00fe Xn i\u00bc1 GiT i \u00fe Xn i\u00bc1 HiRi \u00bc 0; \u00f015\u00de where J is the Jacobian matrix. By extracting the base dynamic parameters from Ri and Ti, Eq. (15) can be rewritten in linear form as JT s Xp \u00bc 0; \u00f016\u00de where X is the observation matrix containing kinematics information, and p is the base dynamic parameters. In the following sections, the proposed method is used to a 3- DOF redundant parallel manipulator and its corresponding nonredundant parallel manipulator. The kinematic model of the 3-DOF parallel manipulator is shown in Fig. 1. Sliders E1D1 and E2D2 drive links A1D1 and A2D2 when they slide along the vertical guide ways. Links E1B1 and E2B2, to be driven by two actuators, are extendible struts. G1 and G2 are counterweights. The whole construction enables movement of the moving platform in a plane and its rotation about the axis normal to the motion plane of the manipulator. Link E1B1 is a redundant link. In this paper, the corresponding non-redundant parallel manipulator is the one shown in Fig. 1 with link E1B1 removed. The kinematic model is shown in Fig. 1. Let the position vector of joint point A1 be rO0 \u00bc x y\u00bd T in the base coordinate system, and the position vector of point Bi (i = 1,2) can be expressed as rBi \u00bc \u00bd xBi yBi T \u00bc rO0 \u00fe Rr0Bi; \u00f017\u00de where R is the rotation matrix from coordinate system O 0 \u2013x 0 y 0 to O\u2014XY; r0 Bi is the position vector of point B1(B2) in O 0 \u2013x 0 y 0 . Based on Fig. 1, the following equations can be obtained: sin /i \u00bc xBi xEi li ; cos /i \u00bc yBi qi li ; i \u00bc 1;2; \u00f018\u00de sin bi \u00bc x xDi l ; cos bi \u00bc y \u00f0qi \u00fe l5\u00de l ; i \u00bc 1;2; \u00f019\u00de where /i is the angle between link EiBi and the vertical axis parallel to the Y-axis, bi is the angle between link AiDi and the vertical axis parallel to the Y-axis, d is the distance between two columns, qi and yBi are the Y coordinates of points Ei and Bi, li is the length of the ith extendible link, l is the length of the constant length link, and xEi, xDi and xBi are the X coordinates of points Ei, Di and Bi in O\u2013XY", + " (18) and (19), the inverse kinematic solutions of the redundant manipulator can be written as q1 \u00bc y l5 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi l2 \u00f0x\u00fe d=2\u00de2 q ; \u00f020a\u00de q2 \u00bc y l5 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi l2 \u00f0x\u00fe d=2\u00de2 q ; \u00f020b\u00de l1 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x l6 sin a\u00fe d=2\u00de2 \u00fe \u00f0y l6 cos a q1\u00de 2 q ; \u00f020c\u00de l2 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x l6 sin a\u00fe d=2\u00de2 \u00fe \u00f0y l6 cos a q2\u00de 2 q ; \u00f020d\u00de where l5 is the height of the slider, a and l6 are the rotation angle and length of the moving platform, respectively. For the configuration shown in Fig. 1, the \u2018\u2018\u00b1\u201d of Eq. (20) should be \u2018 \u2019. For the corresponding non-redundant parallel manipulator, the inverse kinematic solution is shown as Eqs. (20a), (20b) and (20d). Taking the time derivative of Eqs. (18) and (19) leads to _qi \u00bc _y\u00fe l sin bi _bi \u00bc J i\u00bd _x _y _a T; \u00f021\u00de _li \u00bc _xBi sin /i \u00fe \u00f0 _yBi _qi\u00de cos /i \u00bc J i\u00fe2\u00bd _x _y _a T; \u00f022\u00de _bi \u00bc _x l cos bi ; \u00f023\u00de where J i \u00bc \u00bd tan bi 1 0 ;E \u00bc 0 1 1 0 T , J i\u00fe2 \u00bc \u00bd sin /i cos /i eT 1 eT 2 \" # \u00fe ERr0Bie T 3 ! J i cos /i; e1 \u00bc \u00bd1 0 0 T; e2 \u00bc \u00bd 0 1 0 T; e3 \u00bc \u00bd0 0 1 T: Thus, the Jacobian matrix of the redundantly actuated parallel manipulator studied here can be expressed as J \u00bc \u00bd JT 1 JT 2 JT 3 JT 4 T : \u00f024\u00de For the corresponding non-redundant parallel manipulator, the Jacobian matrix can be expressed as JN \u00bc \u00bd JT 1 JT 2 JT 4 T : \u00f025\u00de The base dynamic parameters of the parallel manipulator include mass, the first moment and moment of inertia" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003456_s11071-013-1004-7-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003456_s11071-013-1004-7-Figure3-1.png", + "caption": "Fig. 3 Large deflection of a cantilever beam problem 0 < {\u2212, :}[l \u00d7 b \u00d7 t,E \u00d7 \u03bd]{+,\u00b1}\u2032\u2032(,, \u2212 P )", + "texts": [ + " The solution time was scaled, providing a relative calculation time of 1 for the model having the elastic modulus E = 2 \u00d7 108 Pa. For a very soft material, the use of the solver ode45 gives the best performance, whereas the models with materials that can be used in real-life applications demonstrate better times with solvers ode15s and ode23t. In this section, the results of numerical simulations for different problems are presented. 3.1 Cantilever beam subjected to large bending Two equal vertical forces cause a large displacement of a thin beam (Fig. 3). The length of the beam L = 0.5 m, the width of the beam b = 0.25 m, its thickness t = 0.01 m. Gravity was not accounted for in this solution. The magnitude of the total vertical load P was selected to provide discrete values of the dimensionless force parameter p = PL2/EI , where I is the moment of inertia of the beam\u2019s cross-section.1 1This problem can unambiguously be denoted by text line 0 < {\u2212l/2,\u2212b/2 : \u00b1b/2}[l\u00d7b\u00d7 t,E\u00d7\u03bd]{+l/2,\u00b1b/2}\u2032\u2032(0,0,\u2212P ) for the further external references. This line contains all data mentioned above: geometry and material parameters of the plate are given in the square brackets [" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003035_asjc.547-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003035_asjc.547-Figure2-1.png", + "caption": "Fig. 2. Producing principle of Lorentz radial suspension forces.", + "texts": [ + " The magnetic field produced by the suspension force winding breaks the magnetic field balance established by the torque winding, which makes the flux density in one side gap increase and the flux density in the symmetrical other side gap decrease. The magnetic force, i.e., Maxwell force, is produced toward the direction of the magnetic field increase. Fig. 1 shows the principle of Maxwell force generation of the BPMSM in the x- and y-direction. Besides Maxwell force, the currents of radial suspension force windings are also affected by Lorentz force. Fig. 2 shows the principle of Lorentz force generation of the BPMSM in the x- and y-direction. The radial suspension forces of the BPMSM can be modeled as: \u00a9 2012 John Wiley and Sons Asia Pte Ltd and Chinese Automatic Control Society F K K i i F K K i i x M L Bd Md Bq Mq y M L Bq Md Bd Mq )(= + + = + \u2212 \u23a7 \u23a8 \u23a9 ( ) ( )( ) \u03a8 \u03a8 \u03a8 \u03a8 (1) where K P P L lr W k W k M M B m M WM B WB = \u03c0 \u03bc 2 08 and KL M B WB M WM = 3P W k rW k4 are Maxwell force constant and Lorentz force constants, respectively; Fx and Fy are the radial suspension forces in xand y-direction composed of Maxwell forces and Lorentz forces, respectively; iBd and iBq are current components of suspension force windings in d-q coordinates, respectively; YMd and YMq are the air gap flux linkages components of torque windings and PMs of rotor in d-q coordinates, respectively; l is the length of rotor iron core; r is the radius of the stator inner surface; Lm2 is the mutual inductance of suspension force windings; WM and WB are the number of turns of torque windings and suspension force windings, respectively; and kWM and kWB are winding factor of torque windings and suspension force windings, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003381_20120215-3-at-3016.00075-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003381_20120215-3-at-3016.00075-Figure4-1.png", + "caption": "Fig. 4. Schematic of the pushing V-belt CVT (a) and the geometry of the pulley (b).", + "texts": [ + " The behavior of such a transmission is similar to the basic electric variator transmission, with the following main differences: (i) at low velocities the M/G1 power can be increased to fill the battery and extend the optimal engine operation to lower velocities, (ii) in the mid velocity range the battery can provide additional power for supplying the M/G2 motor for a torque boost (for the same reason, there is no purely mechanical mode); and (iii) in the high-velocity range, the M/G1 is motoring like in the variator mode (Pmg1 < 0), but there is no real electric power recirculation since the M/G2 is motoring, as well, based on the battery power supply. The main parts of the CVT include two V-shape pulleys connected by a metal belt, as illustrated in Fig. 4a. Power is transmitted by compression of the individual segments of the metal belt rather than tension as is done with rubber belts. To change the transmission ratio (the ratio between the secondary pulley radius to the primary pulley radius), the movable halves of each pulley are axially adjusted under hydraulic control. The geometrical quantities of the pulley are shown in Fig. 4b In terms of these quantities, before constructing the bond graph of the pulley/belt, the kinematic relationship between the pulley radius, R, and pulley angle, \u03b1, has to be derived. Then by differentiating this equation, a relationship between the belt radial velocity, vR, and the axial pulley velocity, vP, is obtained. In order to account for frictional losses between the belt and the pulley, the kinematic relationship between the relative speed of the belt with respect to the pulley is derived" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003692_978-3-642-23681-5_13-Figure13.27-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003692_978-3-642-23681-5_13-Figure13.27-1.png", + "caption": "Fig. 13.27 Pressure contours from FLUENT 5.0 for 1,000 rpm, 50 kPa 2976.35 N, 3 mm, 36 groove (Pressure in Pascals) (Pai et al. [25])", + "texts": [ + " \u2022 P8A to P8F measure the pressure along the groove, and exhibit almost constant pressures at values higher than the supply pressure. There is a sharp drop from the pressure in the groove to the inlet and outlet values. \u2022 Power drawn by the electric motor generally increased as the load and speed increased. \u2022 The temperature rise in the water as it passed through the bearing increased as the load increased and was higher for 1,500 rpm than for 1,000 rpm. The result of the CFD analysis is presented in Fig. 13.27. Table 13.2 Theoretical calculations for the 36 groove bearing [16] Load (N) Speed (RPM) Eccentricity ratio Minimum film thickness (mm) Attitude angle (degrees) Fig. 13.26 Circumferential pressure distribution of 3 mm, 36 groove (1,000 rpm, 100 kPa) (Hargreaves et al. [16]) The advantages of axial groove journal bearings with water as a lubricant over less environmentally friendly oil-lubricated bearings are becoming increasingly recognised and understood. In terms of regular use, preventive maintenance and operating costs can be potentially reduced" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003692_978-3-642-23681-5_13-Figure13.1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003692_978-3-642-23681-5_13-Figure13.1-1.png", + "caption": "Fig. 13.1 A schematic diagram of the clearance space of the bearing showing two representative axial groove", + "texts": [ + " Lubricated journal bearings have been widely used since the Egyptian Empire (in chariots lubricated by animal fat) and formed a critical element in the Industrial Revolution. The number of journal bearings currently in use worldwide, well exceeds the population of the earth. The design of journal bearings is considered important to the development of rotating machinery. A journal bearing consists of a shaft completely or partially surrounded by a sleeve. Differences in their respective diameters result in a small clearance between the shaft and the sleeve\u2014see Fig. 13.1. Bearings use a lubricant between the bearing surfaces to drastically reduce the friction force and resultant wear which in turn prevents large temperature gradients. Sliding surface bearings have the common aim of ensuring a good hydrodynamic oil film lubrication when operating, thereby avoiding contact with the shaft. Commonly oil is used, but other liquids such as water can also act as a lubricant. Although the viscosity of water is about 30\u201340 times less than mineral oils, it can be used with appropriate design considerations", + " Some research on the groove arrangement in journal bearings has been conducted [18, 19]. They analysed the performance of journal bearings with oil grooves that were positioned at the maximum pressure location. They discovered that positioning the grooves at the maximum pressure location will cause 30\u201370 percent reduction in the load capacity of the bearing. However these results are not applicable to journal bearings with multiple axial grooves supplied with water from one end of the bearing only (Fig. 13.1). Theoretical calculations have also been undertaken to verify and augment the experimental data generated [16, 17], see Fig. 13.3. The flow in a journal bearing supply groove has an important role in determining the performance of the journal bearing, represented by its load-carrying capacity and energy consumption. The flow in the grooves and land area determines the pressure field along the grooves and over the land area. Estimation of this pressure will give the load capacity (that can be supported by the lubricant film), coefficient of fluid friction and volume rate of flow", + " Some design procedures for axial groove journal bearings have been reported [22, 23] but once again, these do not cater for multiple axial grooves supplied with lubricant from one end only. The configuration of the water lubricated bearing is similar to the submerged bearing studied by Pai and Majumdar [24]. The lubricant (water) is fed from one end through multiple axial grooves. Groove depth is kept at 3 mm but the number and size of grooves may vary. The lubricant film is generated in the non-grooved (land) region (see Fig. 13.1). The flow in this region may be both circumferential and axial. The lubricant flows out from the bearing ends axially [25]. They found that the maximum pressure in the clearance space of the bearing does not occur at the central plane but shifts closer to the outlet side of the bearing. This is because the lubricant is supplied axially. At the inlet, flow into the bearing takes place only in the unloaded region. At the outlet, flow takes place out of the bearing in the loaded region. The flow into and out of the bearing is maintained very much similar to that in a submerged bearing The pressure generated due to wedge action in the clearance space supports the applied load without metal-to-metal (or solidto-solid) contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000826_j.euromechsol.2009.07.005-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000826_j.euromechsol.2009.07.005-Figure3-1.png", + "caption": "Fig. 3. General representative of {X(u)}{X 3(v)} generators of X\u2013X motion.", + "texts": [ + " The equality {T}{X 3(v)}\u00bc {X(v)} implies that {X 3(v)} must not be included in {T}, or else {T}{X 3(v)}\u00bc {T} s {X(v)}. The 1D manifold {X 3(v)} has to include the rotation. Implemented by only lower pairs, an {X 3(v)} manifold cannot but be only a group {H(N, v, p)} with HH(R- )H PH(R- )H HP(R- )H PP(R- )H R R R R a b c d a given pitch that can be any real number. Hence, in general, we can write {X(u)}{X(v)}\u00bc {X(u)}{H(M, v, q)}\u00bc [{H(N1,u, p1)}{H(N2,u, p2)}{H(N3,u, p3)}{H(N4, u, p4)}]{H(M, v, q)} and its corresponding realization is depicted in Fig. 3. The generators of {X(u)}{X(v)} motion, which belong to the category modeled by {X(u)}{X 3(v)}, are the serial concatenations of a generator of {X(u)} motion and an H or R pair with any axis parallel to v. By the same token, one can also establish {X(u)}{X(v)}\u00bc {X 3(u)}{X(v)}. The corresponding generators are the serial concatenations of an H or R pair with any axis parallel to u and a generator of {X(v)} motion. The kinematic inversion of a chain corresponds to the exchange of the fixed body and the moving end body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003673_978-94-007-5006-7-Figure5.2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003673_978-94-007-5006-7-Figure5.2-1.png", + "caption": "Fig. 5.2 External forces and moments on link #k", + "texts": [ + " Chapter 8 will illustrate the dynamic analyses of closed-loop systems using the proposed methodology. Finally, it is to be noted that Eq. (5.12) has similar representation as that of a single-chain serial system (Saha 1999b), however, the elements of the matrices and vectors differ, and this will be evident in Sect. 5.2. The forces and moments due to gravitational acceleration or link-ground interactions are two typical examples of external forces that one needs to take care of, particularly, for the robotic systems analyzed in this book. Figure 5.2 shows the kth link subjected to external forces and moments at the points 1, 2 and 3. The vectors connecting the origin of the link Ok to the points of application of these forces and moments are represented by sk;1, sk;2, and sk;3, respectively. Resultant wrench acting at the origin of the link k is then defined as wF k \" nF k fF k # (5.14) where nF k and fF k are the resultant moment about and the resultant force at Ok for the kth link. If there are nf points of application of the external forces and moments, then wrench wF k is given by wF k D nfX j D1 Sk;j wF k;j ; where Sk;j 1 sk;j 1 O 1 and wF k;j \" nF k;j fF k;j # (5.15) In Eq. (5.15), Sk;j is the 6 6 matrix, and sk;j 1 is the cross-product tensor associated with vector sk;j , for j D 1, 2, 3, as shown in Fig. 5.2. The elements of the GIM, I, obtained in Eq. (5.13) will be derived in this section. These not only enable the factorization of the GIM but are also required for formulating the inverse dynamics algorithm reported in Chap. 7. For that, the GIM is written as I NT d QMNd ; where QM NT l MNl (5.16) In Eq. (5.16), the matrix QM is referred to as the generalized mass matrix of composite modules, and the matrices Nl and Nd are the module-DeNOC matrices. Substituting the expressions for Nl and the mass matrix M from Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001677_s00170-010-2705-4-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001677_s00170-010-2705-4-Figure2-1.png", + "caption": "Fig. 2 Planar force analysis diagram", + "texts": [ + " The analysis emphasises the necessity for determination of the contact state between links rather than establishing formulations for the identification of the conditions that ensure successful height adjustments of a cylindrical pair. 2.1 Force/moment analysis The inner link of the cylindrical lower pair is mounted to a static environment, while the outer link is attached to an articulated manipulator wrist. It should be noted that static does not imply rigid, where the stiffness of the environment is theoretically infinite. The free body diagram for planar cylindrical pair height adjustment under static conditions is depicted in Fig. 2. The friction coefficient \u00b5 is assumed to be identical at all contact points. No stiction effects are considered. Consequently, a distinction between dynamic and static friction is unnecessary. For the convenience of analysis, the height adjustment is assumed to be along the positive Z direction. The clearance between the links is also exaggerated for the purpose of visualisation, and does not represent the true scale of parts. Dynamics due to the rotation of the outer link about its longitudinal centre axis are assumed to be negligible, and the outer link is considered to be exactly symmetrical about the longitudinal axis", + " In such a situation, the information concerning the sense of misalignment \u03b8 is essential for robotic control. In order to identify the sense of the misalignment, estimated forces are utilised to evaluate the moments for both positive and negative misalignments, and the determined moments are subsequently compared with the measured moments. & Contact points A and B are on the negative and positive segments of Xp, respectively, at all times. This implies that the distance between the contact points and point O in Fig. 2 is likely to be lt or lt+lw, depending on the sense of \u03b8. According to the near zero values of \u03b8 that are positive or negative, the moments can be determined by Eq. (5) or (6). These moments are 332 referred to as the calculated moments for the convenience of description. Mc 3 \u00feq\u00f0 \u00de \u00bc falt fb lt \u00fe lw\u00f0 \u00de mR fa fb\u00f0 \u00de \u00f05\u00de Mc 3 q\u00f0 \u00de \u00bc fa lt \u00fe lw\u00f0 \u00de fblt mR fa fb\u00f0 \u00de \u00f06\u00de & Given the fact that the measured moment M3 is equivalent to the calculated moment Mc 3 q\u00f0 \u00de only if Mc 3 q\u00f0 \u00de is calculated with the appropriate sense of \u03b8, the comparison of Mc 3 \u00feq\u00f0 \u00de and Mc 3 q\u00f0 \u00de to the measured moment M3 provides the correct sense of angular misalignment \u03b8, since the calculated moment Mc 3 depends on the location of contact points A and B", + " Subsequently, the determined sense and contact forces fa and fb can be utilised by the force and position controllers of the system to minimise the forces between the links. The use of moment M3 for identification of the sense of the misalignment is unique. Essentially, the process involves the determination of fa and fb, which produces the same Mc 3 \u00feq\u00f0 \u00de and Mc 3 q\u00f0 \u00de failing the orientation identification process. The relevant condition is described as Mc 3 \u00feq\u00f0 \u00de \u00bc Mc 3 q\u00f0 \u00de \u00f07\u00de Substituting Eqs. (5) and (6) into Eq. (7), the contact force relationship can be established fa \u00bc fb \u00f08\u00de & From Fig. 2, it is evident that the above condition would never be satisfied unless sticking is present between the two links, i.e., the links are adhered at the contact point. Therefore, the proposed technique to determine the sense of the misalignment is valid for all values of contact forces. When the links are in a two-point contact situation, the controller action comprises a translational and a rotational correction, which depend on the contact forces and resultant moments. & Line contact Line contact can be considered as a variation of single-point contact, where the contact force is arbitrarily distributed along the line of contact between the links" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure11-1.png", + "caption": "Fig. 11 Arrangement of coordinate systems Sf, Sn, and Sp in case of duplex type of taper", + "texts": [ + " Machine-tool setings are given by S2 = A2 + R2 2AmRu sin 36 ig. 9 Location of the blades and reference point P in coordiate system SP r2 m u ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 03/17/201 tan q2 = Ru cos Am Ru sin 37 Em2 = 0 38 XB2 = \u2212 b2 39 XD2 = 0 40 m2 = R2 41 g1 1 42 g2 2 43 m2c2 = 1 sin 2 44 4.3 Derivation for Duplex Type of Taper. In case of a spiral bevel gear drive with duplex type of taper or the tilted root line type , the dedendum angles 1 and 2 are given by values represented in Table 1. Figure 11 shows the location of systems Sp, Sn, Sf and S2. The root line and the pitch line are not parallel each other and does not intersect at apex of pitch cone, point O2, as in the case of standard type of taper. The distance PPo is given as in the case of standard type of taper by PPo=b2 cos 2. The distance OnO2 is given by OnO2=Am sin 2\u2212 PPo and represents the sliding base. The derivations made for standard type of taper are similar for duplex type of taper. The machine-tool settings are given by OCTOBER 2010, Vol", + " In the particular case of the uniform type of taper, the root line nd the pitch line are parallel each other. The dedendum angle 2 s equal to zero and the gear machine-tool settings can be obtained aking 2 equal to zero at the relations obtained for duplex type f taper. Determination of Point Width The point width affects to the thickness of the gear teeth and to he pitch sharing between the pinion and the gear teeth. Pitch oint P at the gear is used to be located closer to the face line see, or example, Fig. 11 than to the root line. If the pitch would be hared at the pitch cones of pinion and gear on point P, gear teeth ill have larger root thickness than pinion teeth. In order to balnce the thicknesses of pinion and gear teeth, a different point rom point P should be chosen for the pitch to be shared between inion and gear teeth. The proposed approach for determination of he point width is as follows. i An auxiliary coordinate system Sk is defined rigidly connected to the head-cutter see Fig. 12 a . Axis zk is collinear to axis zg whereas axis xk is parallel to axis xg and contains point N" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002614_j.phpro.2012.02.004-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002614_j.phpro.2012.02.004-Figure5-1.png", + "caption": "Figure 5. contact stresses in the outer ring and roller", + "texts": [], + "surrounding_texts": [ + "contact stresses are calculated. The stress distributions calculated are shown by Figs. 5 and 6. Zhang Yongqi et al. / Physics Procedia 24 (2012) 19 \u2013 24 23 Author name / Physics Procedia 00 (2011) 000\u2013000 Figure 7 shows that after the force of inertia taking into account, the contact stress 10% higher than the case without inertia force, so the main gear axle inertial forces at work can not be ignored. 24 Zhang Yongqi et al. / Physics Procedia 24 (2012) 19 \u2013 24 Author name / Physics Procedia 00 (2011) 000\u2013000 Ackonwledge This paper is supported by Graduate Innovation Fund of Jilin University (No.20101024)." + ] + }, + { + "image_filename": "designv11_3_0001189_s10800-008-9578-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001189_s10800-008-9578-3-Figure1-1.png", + "caption": "Fig. 1 Column for granulated carbon potential measurements", + "texts": [ + " Undoubtedly, ensuring a good electrical contact between the granules is an important condition of obtaining reproducible measurements of activated carbon potential. Therefore, it was especially important to provide proper electrical contact between granules, at the same time keeping the pressure on the granules low, so as to avoid destruction of granules and dust formation. The above considerations were taken into account in designing a special column for measurements with granulated carbon, shown on Fig. 1. It was found that tightness, good conductivity, and integrity of carbon granules were ensured if the dry weight of carbon sample was ca. 2.0\u20133.0 g (or the weight of a sample wetted with deionized water, ca. 3.5\u20137.0 g), depending on the brand of activated carbon. The electrical resistance of active carbon samples in the assembled device was 24.8\u2013250 X, while the electrical resistance of wetted samples was 6\u201364 X. The column comprised the casing 1 and clamp cover 2, both made of Teflon, stainless steel bottom 3 and upper 4 meshes that were quite elastic in order to provide the necessary pressure on the carbon granules 5 placed inside an empty compartment in the casing base 1 (Fig. 1). The device was assembled in the following way: a sample of activated carbon was placed into the empty space inside the base 1, then the upper mesh 4 and rubber gasket 6 were placed on top, with the cover 2 placed over it. Four screws 7 with nuts 8 and washers 9 were used to secure the cover 2 to the base 1. In order to avoid having a galvanic pair that would interfere with potential measurements, no metallic Table 1 Brands and characteristics of studied activated carbons Brand Size of granules (mm) Packed density (g/dm3) Mechanical article strength (%) Total pore volume (cm3/g) Micropore volume (cm3/g) Specific surface (m2/g) Raw stock AG-3A 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001117_045104-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001117_045104-Figure6-1.png", + "caption": "Figure 6. Measurement principle of the tooth lead form of an involute helical gear.", + "texts": [ + " One of the most typical measurement methods is the generating method, where the measured gear rotates and the probe moves by rb\u03b8 along the line of action of the measured gear; in other words, the relative motion of the probe generates the theoretical involute curve, and the displacement sensor attached to the probe outputs the difference between the theoretical involute curve and the real gear tooth flank. This measurement method is adopted in this research, although the concept of the VGC is applicable to any measurement method. Assume that the lead of the helical gear is measured at the position rl from the axis of the gear, and the helix angle at this radial position is \u03b2 as shown in figure 6. The theoretical helical curve is as follows, where the axial position is h and the rotational angle of the gear is \u03b8 l: \u03b8l = h tan \u03b2 rl . (3) The lead deviation is the tooth form deviation along the helix in the direction normal to the theoretical tooth flank in the section normal to the gear axis. Some types of the measurement method of the lead deviation have been reported; for example, the probe moves relative to the measured gear according to equation (3), and the displacement sensor attached to the probe outputs the difference between the theoretical helix and real gear tooth" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000183_tmag.2008.2002996-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000183_tmag.2008.2002996-Figure5-1.png", + "caption": "Fig. 5. Cross sections and saturation of the motors at the rated current. (a) Massive rotor with a load angle of 50 . (b): Flux barrier rotor with a load angle of 60 .", + "texts": [ + " After every electrical cycle, the flux linkage of each phase is available as a function of the time. Numerical derivations using spline interpolation provide the EMF waveforms (15) The mean value of the torque is computed by (16) where is the angular velocity of the rotor and the period of an electrical cycle. The power factor is approximated by (17) where is the RMS value of the fundamental of the EMF and I the RMS value of the current. The main parameters of the studied motors are given in Table II. The motors have the same stator. Cross section views of each motor are given in Fig. 5(a) and (b). Fig. 5 also shows the level of saturation in each motor when fed by the rated current and operating at the maximum torque. The corresponding load angles are, respectively, 50 for the massive rotor and 60 for the flux barrier SynRM. V. VALIDATION WITH THE FINITE-ELEMENT METHOD The proposed reluctance network modeling is up to 7 or 9 times faster than the finite-element computation for the flux barrier and the massive rotor machines, respectively (Table III). In order to validate this modeling approach, we compare, for both machines, the EMF waveforms obtained from the reluctance network models with those obtained from the finite-element method [8]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002501_c1ay05344b-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002501_c1ay05344b-Figure1-1.png", + "caption": "Fig. 1 Cyclic voltammograms of FMCNPE at 10 mV s 1 in 0.1 M phosphate buffer (pH 7.0): (a) In the presence and (b) in the absence of 0.3 mM LD; (c) and (d) for an unmodified carbon-paste electrode in the absence (c) and presence (d) of 0.3 mM LD, respectively.", + "texts": [ + " Results showed that the maximum electrocatalytic current for LD was obtained at pH 7.0. Therefore, pH 7.0 was chosen as the optimum pH for the determination of LD at FMCNPE. The utility of the modified electrode for oxidation of LD was evaluated by cyclic voltammetry. The cyclic voltammetric responses of a bare carbon-paste electrode in 0.1 M phosphate Anal. Methods, 2011, 3, 2562\u20132567 | 2563 Pu bl is he d on 0 6 O ct ob er 2 01 1. D ow nl oa de d by U ni ve rs ity o f C hi ca go o n 30 /1 0/ 20 14 1 0: 12 :4 2. buffer (pH 7.0), without and with LD, are shown in Fig. 1 (curves c and d, respectively). Fig. 1a and b show cyclic voltammograms of modified electrode in the buffer solution with 300 mM of LD and without LD, respectively. The results show that the sensor produces a large anodic peak current in the presence of LD without a cathodic counterpart (Fig. 1, curve a). That the current observed is associated with LD oxidation and not the oxidation of modifier is demonstrated by comparing the current in Fig. 1 (curve b, without LD) with the one in the presence of LD in Fig. 1 (curve a). It is apparent that the anodic current associated with the surface-attached materials is significantly less than that obtained in the solution containing LD. At the surface of a bare electrode, LD was oxidized to around 580 mV. As can be seen, the electroactivity of LD on the modified electrode was significant (Fig. 1, curve a), with strongly defined peak potential, around 360 mV vs. Ag/AgCl/KCl (3.0 M) electrode. Thus, a decrease in overpotential and enhancement of peak current for LD oxidation is achieved with the modified electrode. Such behavior is indicative of an EC0 mechanism.39 The effect of scan rate on the electrocatalytic oxidation of 300 mM LD at the modified electrode was investigated by cyclic voltammetry. The oxidation peak potential shifts with increasing scan rates toward a more positive potential, confirming the kinetic limitation of the electrochemical reaction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure1-1.png", + "caption": "Fig. 1. Typical structure of a precise cable.", + "texts": [ + " Section 3 expounds the theoretical derivation of the transmission capability including bending rigidity for the precise cable drive system. Section 4 investigated the affect and sensitivity of major design parameter to the transmission capability. Conclusion is in Section 5. Steel cable is a uniform helical arrangement of wires concentrically stranded together for a variety of operating conditions. A typical steel cable consists of several strands preformed around a core or center strand [15]. Each strand includes several wires. Typical structural is shown in Fig. 1. The cable construction has more wires and consequently more flexible with less abrasion resistance. It is difficult to determine the bending stiffness of steel cable. On one hand, elastic modulus of steel cable is not a constant value. It relies on the constitution, and varied with the working time, external load, and load changing frequency. On the other hand, the moment of inertia of steel cable cannot be expressed with a certain formula, since it not only depended on cable radius, but also the structure, the number of wires, the core and so on" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003219_s00170-011-3664-0-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003219_s00170-011-3664-0-Figure2-1.png", + "caption": "Fig. 2 Schematic of a specimen for measuring scanning line width and depth", + "texts": [ + " E \u00bc PL VS HS \u00f01\u00de Here E Energy density on surface (J/mm2) PL Laser average power (W) Green specimen Slurry Composition VS Laser scanning speed (mm/s) HS Hatch space (mm) Selecting a high energy density laser scanning to achieve a consolidation mechanism of full melting is the principle of building a fusion part, while the consolidation mechanism of a sintering part is based on the LPS, which is induced with a suitable energy density. Yen et al. [9] investigates laser scanning parameters on the fabrication of ceramic parts by LPS. The hatch space (HS) is based on the overlap ofWL and the layer thickness is based on the overlap of DL, respectively. Figure 2 illustrates the schematic of the specimen for observing WL and DL. WL and DL were measured by an optical microscope equipped with an X-Y platform. For a specific hatch space, scanning energy density was varied with laser power and scanning speed. The consolidation mechanism was verified by observing the topography with a scanning electron microscope. Firstly, high energy density was employed to fabricate the base and the spacer of the specimen. Secondly, a single observing fusion layer was built with slurry A and a specified scanning parameter combination (PL=19 W, VS=60 mm/s and HS=0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001900_2041305x10394408-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001900_2041305x10394408-Figure1-1.png", + "caption": "Fig. 1 Pad schematics", + "texts": [ + " The objective of this study is to analyse the effects of deformation, rotational speed, and load and oil viscosity grade on lubrication performance of tilting pad thrust bearing of a hydraulic turbine. For this purpose, the Reynolds equation, the viscosity\u2013 temperature and density\u2013temperature characteristics of the lubricants, film thickness equation, energy equation, and heat conduction equation were simultaneously solved by the finite-difference method and a successive over-relaxation method was adopted to obtain the solutions of the above equations. 2 MATHEMATICAL MODEL Figure 1 shows the schematics of bearing pad. The assumptions of the mathematical model are made as follows: the bearing is operating in the steady state; and there is no boundary slip at the fluid\u2013solid interface. The lubricant is a Newtonian fluid; its viscosity and density are functions of temperature only. The inertia and body force terms are negligible compared to the viscous and pressure terms in the momentum equations. The lubricant flow remains laminar and the pressure of the oil film is constant across the film" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003478_iccas.2013.6704123-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003478_iccas.2013.6704123-Figure14-1.png", + "caption": "Fig. 14. Start and Goal position and direction in an experiment", + "texts": [], + "surrounding_texts": [ + "In this paper, we applied the optimization method to the path planning for the wheel loader on scooping and loading operation. We showed the quasi-optimized path and demonstrated by the miniature wheel loader robot. Owing to this method, we could generate shorter path than the method for V-shape [8][9]. Now, we tackle to confirm our method from theoretical view." + ] + }, + { + "image_filename": "designv11_3_0002804_tro.2011.2181098-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002804_tro.2011.2181098-Figure1-1.png", + "caption": "Fig. 1. Dynamic nonprehensile manipulation to rotate (a) a rigid body and (b) a deformable body. The object\u2019s deformation generated by dynamic effects, as shown in (b), can decrease the negative moment. As a result, the object can rotate faster.", + "texts": [ + " These works done on manipulation utilizing a plate have supposed that the object is a particle(s) or a rigid body(ies) since it is convenient from the viewpoint of both geometric and dynamic analyses. In contrast, conventional works treating a deformable object have generally supposed two or more fingers to grasp and handle the object [16]\u2013[23]. As far as we know, there are no works dealing with a deformable object as the main target in nonprehensile manipulation. 1552-3098/$31.00 \u00a9 2012 IEEE Our former works treated a dynamic manipulation that is inspired by the handling of a pizza peel [24], as shown in Fig. 1(a). A chef handles the peel and remotely manipulates a pizza on the plate. We found that the chef aggressively utilizes two DOFs from the remote-handling location during the manipulation: translation X along the bar and rotation \u0398 around the bar. We proposed a dynamic nonprehensile manipulation to control the position and the orientation of a rigid object on a plate by applying the peel mechanism to the robot system. This manipulation scheme has the advantage that the robot can remotely manipulate an object in areas with high temperature, high humidity, electromagnetic field, etc., where a gripper or a robot hand with electronic/electrical/precision device(s) is unavailable. Furthermore, since the object is not grasped or picked, the concentration of stress inducing the object\u2019s destruction is avoided. We have also found that a deformable object can rotate faster than a rigid one [25], [26], as shown in Fig. 1(a) and (b). In order to extend the aforementioned investigation, in this paper we clarify what actually happens in the manipulation for rotating deformable objects. This paper explores the optimal conditions which include the plate\u2019s motion, the object\u2019s physical properties, and the friction between them. In this paper, after showing the principle to rotate an object and some basic experiments, we introduce a simulation model in order to approximate the dynamic characteristics of a thin deformable object on the plate", + " In Section II, we explain the essence of the principle of rotation and show a basic experiment. In Section III, we introduce a simulation model for a deformable object. In Section IV, we show how to estimate the viscoelastic parameters of the object. In Section V, we show the simulation analysis that is based on real food. In Section VI, we give the conclusion of this work. Fig. 2 shows the top view and the side view of the object on the plate; this plate has two DOFs: translation X along the bar and rotation \u0398 around the bar, as shown in Fig. 1(a). The object as well as the plate is stationary (X = 0, \u0398 = 0) in Fig. 2(a). Then, as shown in Fig. 2(b), by giving a translational acceleration X\u0308 to the plate, an inertial force and a frictional force are generated. In this case, the nominal pressure distribution on the object is assumed to be uniform, and as a result the frictional force distribution is also uniform, as shown in Fig. 2(b), where we illustrate just the slice of the frictional force distribution that passes through the center of mass of the object", + " In Fig. 6, it can be seen that the object is bent even more than in Fig. 5, decreasing the contact area between the plate and the object to almost a half, decreasing the braking moment and, thus, rotating faster. In this case, the object rotates with an angular velocity of 251.7 \u25e6/s. In preparation for the motion analysis, we introduce a viscoelastic model to approximate dynamic behaviors of a thin deformable object on a plate. Assumptions: Consider a plate and a thin deformable object as shown in Fig. 1(b). To simplify the analysis, we set the following assumptions. 1) The plate is rigid. 2) The plate\u2019s surface area is larger than that of the object. 3) The object is deformable, and its thickness is small. 4) The object is isotropic, and it has uniform mass distribu- tion and uniform viscoelasticity. 5) The nominal pressure distribution on the object is uniform. 6) The friction coefficient between the plate and the object that is based on Coulomb\u2019s law is uniform and is given by \u03bcs and \u03bck for static and dynamic coefficients, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001656_s11431-010-0064-x-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001656_s11431-010-0064-x-Figure10-1.png", + "caption": "Figure 10 Generated tool path by interpolating cutter orientations. (a) Cutter center points; (b) interpolated cutter orientations.", + "texts": [ + " For sake of clarity, only 52 cutter orientations in each row are shown. It can be found that the cutter orientations are smoothed along both the feed direction and pick feed direction. All cutter orientations of the tool path can be generated by quaternion interpolation [24] based on the cutter orientations of mesh points. The machining tolerance of the blade is set as 0.01 mm and the step distance between two rows of the tool path is set as 0.5 mm. 42022 cutter center points are then obtained for the finish machining of the blade, as shown in Figure 10(a). The finishing tool path is generated by interpolating cutter orientations. Some discrete cutters shown in Figure 10(b) describe the tool path. The tool path illustrates that the cutter orientations are wholly smooth both along the feed direction and pick-feed direction. The computational efficiency is also improved by this mesh-based algorithm. The 386.20 s will be spent in computing accessibility cones of the 42022 cutter center points. The feed velocities are simulated to show the validity of the proposed cutter orientation smoothness algorithm. The blade will be machined by a five-axis machine tool with A-Table & B-Head structure", + " The right row of tool path is used to analyze the smoothness of cutting angles along the feed direction. Figure 12 shows the comparison between initial tool path and optimized tool path. It can be observed that the changes of cutting angles are obviously decreased by smoothing cutter orientations. Since the cutting force can be smoothed by decreasing the change of cutting angles, cutting process will be \u201ctuned\u201d by the developed cutter orientation optimization method. The variation of cutter angles along the pick-feed direction can also be analyzed based on the tool path in Figure 10(b). The comparison is shown in Figure 13. The ranges of lead angle and tilt angle are 90\u00b0 and 40.88\u00b0 before cutter orientations are smoothed. By wholly optimizing cutter orientations, the ranges are decreased to 0.53\u00b0 and 2.58\u00b0. The result shows that the proposed algorithm can smooth the cutter angles along the pick-feed direction. This is helpful to improve the machined surface quality. With the orientation-smooth tool path generated by the proposed method, the cutting experiment is performed in a five-axis machine tool with a A-Table & B-Head structure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003110_s12239-012-0022-7-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003110_s12239-012-0022-7-Figure2-1.png", + "caption": "Figure 2. Two adjacent rigid and flexible bodies.", + "texts": [ + " M is the mass matrix, and Q is the force vector, which includes the external forces in the Cartesian space. 2.2. MFBD Formulation The equations for the motion for the rigid body can be expanded from Equation (14) as follows: (15) where the superscript r denotes a rigid body quantity, rr denotes a relative quantity between rigid bodies, and er represents a relative quantity between a flexible body node and a rigid body. A flexible body node is a node where flexion is permitted in a flexible body. The schematic diagram for two adjacent rigid and flexible bodies is shown in Figure 2. The constraint equations between rigid bodies are expressed as a function of the rigid body generalized coordinates qr as follows: (16) Similarly, we can derive the equations of motion for the flexible body as follows: (17) where qe is the generalized coordinates for the flexible body nodes. The superscript e denotes a quantity describing a flexible body node, ee represents a relative quantity between flexible body nodes, and er denotes a relative quantity between a flexible body node and a rigid body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure14-1.png", + "caption": "Fig. 14 Balanced four-bar mechanism by combination of a CRCM-balanced double and a CRCM-balanced single pendulum", + "texts": [ + " Berestov 10 showed a planar four-bar mechanism balanced by RCMs driven by inner gears. This mechanism can be regarded A O l * 2 l * 1 l 1 l * 1 l 2 l * 2 m 2 l 1 l 2 ig. 13 1DOF crank-slider mechanism without CRCMs obained by restricting the motion of the end point of the 2DOF arallel manipulator s a combination of a balanced single and a balanced double pen- ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201 dulum, shown with chain driven CRCMs in Fig. 14. Also for four-bar mechanisms, the different CRCM configurations are applicable just as the substitution of the well-known kinematic relations into the inertia equations of the double and single pendula to obtain the inertia about one of the links. This means that with the equations for the double and single pendula and the kinematic relations the inertia of any four-bar mechanism can be written down easily. It is a special case for which the four-bar linkage becomes a parallelogram. From Fig. 9, and assuming the link between O and A to be fixed with the base, the resulting parallelogram can be balanced, as in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002784_icsengt.2012.6339294-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002784_icsengt.2012.6339294-Figure3-1.png", + "caption": "Fig. 3 DATCOM+ geometric view of the fixed-wing UAV", + "texts": [ + " This paper presents the implementation of both methods to build an empirical linear state-space model based on flight test data. Further, the empirical model obtained from flight test data is used for validating the analytical model by comparing the simulation results of the analytical and empirical models in similar flight condition. If the analytical model is comparable to the empirical model, then it can be used for design of the flight control system of the UAV. The parameters of BPPT01A-PA6 Wulung UAV are listed in Table I, while its geometrical configuration is as shown in Fig. 3. The parameters are required for formulating the analytical model of UAV longitudinal dynamic, and its geometrical details are required for computing the aerodynamic coefficients using DATCOM. DATCOM software computes aerodynamic coefficients required by the longitudinal dynamic model formulation. The computation is performed over some steady state flight conditions, which are determined by some flight parameters values as shown in Table II. The aerodynamic coefficients computed by DATCOM, which are required by the linear model, are listed in Table III" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002606_s00707-012-0799-5-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002606_s00707-012-0799-5-Figure1-1.png", + "caption": "Fig. 1 A nonholonomic mobile manipulator", + "texts": [ + " Finally, a number of simulations and experiments are performed for a spatial nonholonomic mobile manipulator which demonstrates the efficiency and applicability of the proposed method. In this section, the set of nonlinear dynamic equations of the nonholonomic mobile manipulator is presented. Then, the problem of trajectory optimization of the system is formulated which results in the proper solution. 2.1 Kinematic and dynamic model of the system This section presents a general formulation for kinematic and dynamic modeling of the nonholonomic mobile manipulators. The model is composed of robotic arms mounted on a nonholonomic mobile platform as depicted in Fig. 1. To develop the mathematical model of the system, at first, the kinematic constraints of the mobile robot are presented. As it is shown in Fig. 1, the mobile manipulator consists of a platform driven by two independently driving wheels. The nonholonomic constraints of the mobile platform can be stated as: the mobile platform can only move in the direction of the symmetry axis, and each driving wheel must have pure rolling without any slippage. By defining the five-dimensional Lagrangian coordinates of the mobile platform of the system as qw = [ x y \u03c6 \u03b8r \u03b8l ]T, the kinematic constraints of the nonholonomic mobile base can be presented as [21]: Aw(qw)q\u0307w = O, (1) where O is the zero matrix and Aw is the Jacobian matrix of the mobile base. Considering a manipulator with n degrees of freedom mounted on a mobile base (Fig. 1), the generalized coordinate of the system can be defined as qt = [ qw qm ] , where qm = [ \u03b81 \u03b82 \u00b7 \u00b7 \u00b7 \u03b8n ] denotes the n-dimensional Lagrangian coordinates of the manipulator. Thus, the nonholonomic constraints of the mobile manipulator can be stated as follows: A(qt) q\u0307t = O (2) where the Jacobian matrix A is A(qt) = [ Aw O ] 3\u2217(n+5) . Now, by obtaining the kinematic and potential energies of the system and using the Lagrange multiplier principle, the dynamic model of system is described as [22]: Mm(qm)q\u0308m + Vm1(qm, q\u0307m) + Vm2(qm, q\u0307m, ) = \u03c4m \u2212 Rm(qm, qw)q\u0308w (3) Mw1(qw)q\u0308w + Vw1(qw, q\u0307w) + Vw2(qm, q\u0307m, qw, q\u0307w) = Ew\u03c4w \u2212 AT\u03bb \u2212 Mw2(qm, qw)q\u0308w \u2212 Rw(qm, qw)q\u0308m (4) where Mm is the inertia matrix of the manipulator, Vm1 denotes velocity-dependent terms of the manipulator, Vm2 denotes Coriolis and centrifugal terms caused by the angular motion of the platform, \u03c4m is the input torque for the manipulator, and Rm is related to the effect of the mobile base dynamics on the manipulator" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000605_jf803962b-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000605_jf803962b-Figure2-1.png", + "caption": "Figure 2. Substrate concentration profile within the spherical biocatalyst of average radius R and liquid film adhering to the external surface of the biocatalyst itself.", + "texts": [ + " In these circumstances, the urea degradation rate referred to the unit volume of immobilized acid urease (rSi) may be expressed using the following modified form of eq 3 rSi \u00bc kIiS \u00f015\u00de with kIi \u00bc k 0IiFBYP=B \u00f016\u00de where kIi is the urea degradation pseudo-first-order kinetic rate constant of the biocatalyst of concern, FB is the biocatalyst density, YP/B is the protein loading, k0Ii is the specific pseudo-first-order kinetic rate constant relative to immobilized enzyme, and S is the urea concentration in the liquid infiltrating into the bead pores. When using a perfectly mixed bioreactor, charged with a volume (VL) of the model wine solution with an initial concentration of urea SL0 and inoculated with a prefixed concentration (cBd) of dry biocatalyst in the form of almost spherical beads with an average radiusR and specific surface per unit volume (ap), the urea concentration profile within the spherical biocatalyst and through the liquid film adhering the external surface of the biocatalyst itself is sketched in Figure 2. In these circumstances, the overall surface (aS) and volume (vS) for the biocatalyst per unit volume of liquid phase may be estimated as aS \u00bc apcBd=FB \u00f017\u00de vS \u00bc cBd=FB \u00f018\u00de The unsteady-state material balance for urea may be written as SLVL t \u00bc SLVL t\u00fedt \u00fe kLaSVL\u00f0SL -SR\u00de dtjj \u00f019\u00de with the boundary condition kLaSVL\u00f0SL -SR\u00de \u00bc \u03a9\u00f0kIiSL\u00devSVL \u00f020\u00de with \u03a9 \u00bc \u03b7 1 \u00fe \u03b7\u03a62 3Bi \u00f021\u00de \u03b7 \u00bc 3 \u03a6 1 tan h\u00f0\u03a6\u00de - 1 \u03a6 \u00f022\u00de \u03a6 \u00bc R ffiffiffiffiffiffiffi kIi DSe s \u00f023\u00de Bi \u00bc kLR DSe \u00f024\u00de where \u03a9 or \u03b7 is the effectiveness factor for a spherical biocatalyst in the presence or absence of the external film transport resistance, \u03a6 the Thiele modulus for pseudo-firstorder kinetics, Bi the Biot number, which measures the ratio between the external film transport and intraparticle diffusion rates of the reagent of concern, kL andDSe are the mass transfer coefficient in the liquid phase and effective diffusion coefficient for urea, and SR is the reagent concentration at the biocatalyst surface (28 )" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002865_1.3555006-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002865_1.3555006-Figure9-1.png", + "caption": "Fig . 9 F ig . 10", + "texts": [ + "asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jotre9/28554/ on 03/02/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use at higher frequencies. Higher values of j3co0 result in higher rising frequency, while higher values of a mean higher rising rate. On the other hand, for bearings compensated by restrictors having 0 = 1, higher values of a mean higher minimum stiffness and higher values of ffu\u201e result in an increase in the frequency at which this minimum stiffness occurs. Fig. 9 shows, in table form, the effect of the different bearing and fluid design parameters on the change of the dynamic stiffness with frequency. Experimental Investigation Experimental Bearings. Two types of bearings were the subject of the experimental investigation: 1 A single circular bearing, Fig. 10, having a recess radius n = 1.25 in. and an outside radius r2 = 4.00 in. with a rod in its rectangular bearings 1 in. X 2 in. with a 3/8-in. landing. The design parameters for such a bearing were as follows: p s = 750 psi, pb/v\u201e -\u2022 = 0", + " In order to increase the value of b, the free-air percentage should be decreased and this could be achieved by good sump design and other mechanical precautions. However, especially for open hydraulic systems, such as for hydrostatic bearings, these precautions will only decrease the percentage of free air in the system and will not eliminate it absolutely. Accordingly, in order to increase the value of b for such systems, it is essential to avoid low oil pressures. The effect of decreasing the value of K on the dynamic stiffness of hydrostatic bearings is shown in Fig. 9. For bearings compensated by controlled restrictors the effect will be to decrease the minimum stiffness as well as the frequency at which this minimum stiffness occurs. Fig. 17 shows the results of applying a sudden load to the circular bearing previously described. In this figure, the bearing velocity and pressure as well as the applied load are recorded at a speed of 2 psi for different dynamic and static loads. Table 2 shows the average values of K calculated from the test results in Fig. 17" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000648_j.jcat.2007.05.006-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000648_j.jcat.2007.05.006-Figure6-1.png", + "caption": "Fig. 6. Series of real-time SEIRA spectra collected at Ru nanofilm electrode/ 0.1 M HClO4 + 20 mM NaNO2 solution interface sequentially during a potential cycling sweep from 0.9 V to 0.0 V and then back to 0.9 V at 20 mV s\u22121. The time resolution used and shown is 0.4 s. (a) The reference spectrum was taken at 0.0 V. (b) The reference spectrum was the first spectrum at 0.9 V.", + "texts": [ + " High mobility is assumed to also apply to the \u03bd2(O)\u2013NO species on Ru, ensuring a rather uniform dilution of local \u03bd2(O)\u2013NO species concentration in the process of reduction. The dilution of local coverage of \u03bd2(O)\u2013NO species diminishes the dipole\u2013dipole coupling effect and thus further decreases the \u03bdN\u2013O frequency [56,57]. The rapid concurrent movement of the \u03bd2(O)\u2013NO species facilitates the reduction of \u03bd2(O)\u2013NO species occurring randomly throughout the layer and the rapid concurrent lowering of the coverage of NO, contributing to the larger slope of Y2. Fig. 6 shows series of SEIRA spectra collected during the initial cathodic potential scan from 0.9 to 0.0 V and then in the anodic potential scan back to 0.9 V at 20 mV s\u22121 on a Ru nanofilm electrode in 0.1 M HClO4 containing 20 mM NaNO2. The reference spectrum was taken at 0.0 V for Fig. 6a, in which the band located at 1886\u20131860 cm\u22121 in the cathodic scan or 1850\u20131870 cm\u22121 in the anodic scan can be assigned to the \u03bd2(O)\u2013NO species [32,33]. The slight change in band position in the cathodic and anodic scans can be explained as a decrease in surface coverage in the latter. The other band located at 1530\u20131578 cm\u22121 in the cathodic scan or at 1520\u20131576 cm\u22121 in the anodic scan can be assigned to multicoordinated NO coadsorbed with oxygen-containing species [designated \u03bd1(O)\u2013NO species] for it is accrete with \u03bd2(O)\u2013NO species [33]. A very weak and downward-directed band at 1630 cm\u22121 is assigned to \u03b4HOH of interface water at 0.0 V [45], as also detected in case I. Interestingly, a downward-directed band appeared at 1740\u20131750 cm\u22121 in the cathodic scan or at 1740\u20131800 cm\u22121 with a bipolar shape in the anodic scan. This band clearly was not detected in case I, suggesting that a new surface species was produced at lower potentials in case II. To remove the bipolar band feature (related to the new species) in Fig. 6a for easy spectral analysis, the initial spectrum collected at 0.9 V was used as the reference to reconstitute the potential dependent series of spectra, as shown in Fig. 6b. In fact, a unipolar band at 1740\u20131790 cm\u22121 can be clearly identified. Careful examination reveals that this band is present in the low potential region from ca. 0.3 to 0.0 V in the cathodic scan and to ca. 0.7 V in the anodic scan. None of the following nitrogen-containing species may contribute to this band: HNO2 (1670, 1275 cm\u22121) [20], N2O (2227, 1300 cm\u22121), NH+ 4 (3040, 1680, 3145, 1400 cm\u22121), NH2 (matrix: 1500, 3220 cm\u22121; surface: 3290, 1610, 3380 cm\u22121), [NH\u2212 2 ] (3270, 1556, 3323 cm\u22121), HNO (matrix: 3450, 1110, 1563 cm\u22121), or NH3 (3223, 1060, 3378, 1646 cm\u22121) [58]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000166_ls.65-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000166_ls.65-Figure3-1.png", + "caption": "Figure 3. Grid of the bearing liner with eight-node hexahedral elements (N\u03b8 = 30, Nr = 2 and Nz = 10).", + "texts": [ + " Validation of the (3-D) Thin Elastic Liner Model To establish the validity of the (3-D) thin elastic liner model in the case of a compliant fi nite length journal bearing (R/L = 0.5), we developed a separate computer program of structural analysis using Copyright \u00a9 2008 John Wiley & Sons, Ltd. Lubrication Science 2008; 20: 241\u2013268 DOI: 10.1002/ls 3-D FEM. The fi eld of pressure applied to the surface of the journal bearing is calculated starting from the resolution of the zero-order Newtonian Reynolds equation (6) using the fi nite difference method for a hydrodynamic fi nite journal bearing operating at eccentricity ratio \u03b50 = 0.90 (Figure 2). As shown in Figure 3, the bearing liner is discretised with eight-node hexahedral isoparametric elements. It has been divided into 30 elements in circumferential direction (N\u03b8), 10 elements in the axial direction (Nz) and two elements in the radial direction (Nr). In Figure 4, we compare the radial displacement fi eld at the fl uid\u2013liner interface calculated in the mid-plane section of the journal bearing without fl uid\u2013structure coupling for two values of thickness of the thin elastic liner th = 0.5 mm and th = 1.0 mm, and a compressible material (E = 126 GPa, \u03bd = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003618_s11431-013-5291-5-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003618_s11431-013-5291-5-Figure2-1.png", + "caption": "Figure 2 Plane sketch of bearing inner ring.", + "texts": [ + " The failed inner ring was cut from the spalling zone along the axial direction, and the micro-section below the failed raceway surface was presented in Figure 1(b), where it is clear that two subsurface cracks are approximately parallel to the spalling surface along the axial direction. Thus, the distributions of multiple subsurface cracks are used to guide the study on their effect on RCF. In this paper, the inner ring of a 6208 (bearing designation, SKF) deep groove ball bearing is considered as the RCF research object. The main geometric data are: bearing bore diameter d=40 mm, bearing width b=18 mm, steel ball diameter Dw=12.7 mm, diameter of ball\u2019s centre in the bearing Dpw=60 mm, inner raceway groove curvature radius ri =6.54 mm, as shown in Figure 2. Subsurface cracks are modeled as three elliptical pores and located at an approximately depth H (0.2 mm) below the raceway surface. The major size 2a, the minor size 2b and the thickness 2c of these cracks are 0.5, 0.25 and 0.25 mm, respectively, as shown in Figure 3. Symbol \u03b8 indicates the rotation angle of the inner ring. The rotation angle is defined to be positive when the contact position is to the right of the cracks. The moment M is exerted on the inner ring in the anticlockwise direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001181_1.2983146-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001181_1.2983146-Figure9-1.png", + "caption": "Fig. 9. Geometry used to calculate the swing of a bat, showing the position of the bat and the direction of the force on the handle near the start of the swing. In this position is taken to be positive and is taken to be negative. As the bat rotates in a counterclockwise sense, decreases and increases to a positive value when the barrel is above the x axis at a rate =d /dt.", + "texts": [ + " If the bat were allowed to rotate at such a high speed, the handle would push firmly on the batter\u2019s left hand and tend to pull out of his right hand. The reaction force exerted by the batter is such that the left hand pushes on the handle and the right hand pulls on the handle, thereby generating the large negative couple that restricts the total torque on the bat to a value less than 6 Nm throughout the swing. The model used to calculate the swing of a bat in a twodimensional x ,y plane is shown in Fig. 9. Bats are not normally swung in a horizontal plane, but tend to be swung in a plane inclined to the horizontal to project the ball upward or downward. The angle is the angle between the applied force F and the longitudinal axis of the bat, and the angle is the angle between the longitudinal axis of the bat and the x axis. The positive direction of is defined as shown in Fig. 9. When is positive, the torque due to F acts in the \u201cwrong\u201d direction, meaning that the batter needs to exert a positive couple to rotate the bat in the \u201cright\u201d coun- terclockwise direction. The equations of motion are 40Rod Cross icense or copyright; see http://ajp.aapt.org/authors/copyright_permission d2x dt2 = \u2212 F/M cos \u2212 , 2 d2y dt2 = F/M sin \u2212 , 3 d2 dt2 = C \u2212 Fd sin /Icm, 4 where d is the distance between the bat center of mass and the point of application of F on the handle, taken to be 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002257_j.jsv.2011.06.018-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002257_j.jsv.2011.06.018-Figure2-1.png", + "caption": "Fig. 2. Contact stresses: contact spring is assumed to be in line with er, where er corresponds to the radial direction of the ball.", + "texts": [ + " It is subject to the concentrated inertia of the ball at one end and the free boundary condition at the other end. Accordingly, the transverse displacement and its slope of the center of the ball are equivalent to those of the one end of the beam. The ball is in contact with the hemispherical shell-liner inserted into the rigid socket where the liner is modeled as contact stiffness (kc). The pre-axial load (No) and the rotation of the beam generate normal and friction stresses on the contact surface between the ball and the liner as shown in Fig. 2. In this paper, the dynamic stability criteria at the sliding state are investigated for the prediction of squeak propensity. Therefore, the sliding configuration is found under the action of pre-normal and friction loading and then the perturbation at the sliding equilibrium is described as shown in Fig. 3. According to the above scenario, the sliding equilibrium is to be determined first. Under the sliding state, the center of the ball is located at \u2018\u2018A\u2019\u2019 and the corresponding contact location can be described as req c \u00bc xA\u00feRer , (1) xA XAio\u00feYAjo\u00feZAko, (2) where xA is the position vector of \u2018\u2018A\u2019\u2019 at equilibrium and er is the radial direction vector of the spherical coordinates as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003439_j.molcatb.2013.07.003-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003439_j.molcatb.2013.07.003-Figure2-1.png", + "caption": "Fig. 2. Exploded view illustration of the developed cofactor regeneration module.", + "texts": [ + " Samples were repared as described above. . Results .1. Electrochemical cofactor regeneration module PTFE frame units and glass coverings were chosen because of heir excellent chemical inertness and dielectric properties. Anode nd cathode compartment were separated by means of a Nafion embrane, a synthetic polymer with exceptional thermal, chemcal, and mechanical stability, which is commonly employed in lectrochemical devices [19]. Electrode dimensions were chosen imilar to conventional electrochemical lab-scale reactors [20] Fig. 2). RVC was utilized as electrode material as it has a very low chemical reactivity over a broad range of possible reaction conditions regarding solutes, temperatures, and pH values. Other advantages of RVC constitute the very low pressure drop of flowing streams, low material costs, and easy handling [12], which are essential for an all-purpose and scalable electrochemical cofactor regeneration system. Due to the three-dimensional structure of the microporous material, RVC provides very high electrode surface area to volume ratios" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002352_asjc.345-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002352_asjc.345-Figure2-1.png", + "caption": "Fig. 2. The schematics of the PVTOL aircraft dynamics.", + "texts": [ + ", thrust directed to the bottom of the aircraft, 2011 John Wiley and Sons Asia Pte Ltd and Chinese Automatic Control Society Asian Journal of Control, Vol. 14, No. 2, p . 439\u2013453, March 2 P.-C. Chen: Smooth Switching Control with Application to V/STOL Aircraft one can have simplified V/STOL aircraft dynamics which describe the motion of the aircraft in the vertical-lateral direction, that is, considered as a planar vertical/short takeoff and landing (PVTOL) aircraft. The schematic dynamics for the PVTOL are shown in Fig. 2. The aircraft states are the position of center of mass, (X,Y ), the roll angle , and the corresponding velocities, (X\u0307, Y\u0307 , \u0307). The control input is the thrust directed to the bottom of aircraft Ut and the moment around the aircraft center of mass Um . In the case that the bleed air from the reaction control valves or ducts produces force, which is not perpendicular to the pitch axis, there will be a coupling effect between the angle rolling moment and lateral moving force. Let the ratio of lateral force induced by the rolling moment be denoted by \u03b5o, then the aircraft dynamics as shown in Fig. 2 can be written as,\u23a7\u23aa\u23a8 \u23aa\u23a9 \u2212mX\u0308 = \u2212 sin Ut +\u03b50 cos Um \u2212mY\u0308 = cos Ut +\u03b50 sin Um\u2212mg J \u0308 = Um, (34) where mg is the gravity force imposed on the aircraft center of mass and J is the moment of inertia around the axis through the aircraft center of mass and along the fuselage. To simplify the notation of the PVTOL aircraft dynamics (34), the first and second equations in (34) are divided by mg, and the third one by J . Let x :=\u2212X/g, y :=\u2212Y/g, u1 :=Ut/mg, u2 :=Um/J , \u03b5 :=\u03b5oJ/mg, the normalized PVTOL aircraft dynamics are then obtained as,\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 x\u0308 = \u2212 sin u1+\u03b5 cos u2 y\u0308 = cos u1+\u03b5 sin u2\u22121 \u0308 = u2 (35) The term \u201c\u22121\u201d denotes the normalized gravity acceleration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001557_med.2009.5164716-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001557_med.2009.5164716-Figure3-1.png", + "caption": "Fig. 3. The geometric relations of one rotor seen from the center of the quadrotor\u2019s coordinate system.", + "texts": [ + " Experiments support this results because they show that the ideal autorotation starts when / 0 is approximately 1.79. The term 0 is the induced speed when the rotor is hovering. Blade elemental theory is used to calculate forces and torques acting on a rotor by studying a small rotor blade element. The rotor blade is modeled as an airplane wing so that the airfoil theory can be applied. The basics of Blade elemental theory can be found in [4] and are therefore not the focus of this paper. This paper will show how to derive forces and moments for the specific geometric construction of a quadrotor. Fig. 3 shows the relative position of one rotor as it is seen from quadrotor\u2019s body frame. This rotor is placed at a distance D from quadrotor\u2019s body frame origin and forms an angle of ( 45 6 with quadrotor\u2019s body frame x axis. Rotor spins and therefore has an angular speed \u2126. A cross section of a blade elemental \u03948 at a distance 8 from the rotor center is given in Fig. 2. This figure shows the blade element in quadrotor climb mode when there is no lateral movement of the quadrotor. For better clarity the angles are drawn larger than they actually are. Quadrotor\u2019s climb and induced speeds, and respectively, tend to reduce the mechanical angle of attack 9: 8 to the effective angle of attack 9;< 8 . We also have to consider linear blade twisting 9: 8 \u03980 ( > ?\u0398@A . Where \u03980 is the mechanical angle at the root of a blade, and \u0398@A is the linear twist angle. Fig 1 and Fig 3 show the lateral airstream which is produced when rotorcraft moves in horizontal XY plane. Generally speaking, as can be seen in Fig 1, the stream consists of a vertical flow B and a horizontal flow which form the total flow C $ B . We continue with an observation of a small rotor blade element \u0394r. It produces elemental lift and drag force, EF E> and EE> respectively. The total rotor lift is derived by summing all the elemental lifts of all the blade elements. Because of the blade rotation, forces produced by blade elements tend to change both in size and direction", + " In case of torque equations the angles between the forces and directions are easily derived from basic geometric relations which results in the elemental torque equations: Er% E> ( sEkE> s U* \u00b7 ,lm Ut ( 5 6W ( 8W (8) Er , E> ( \u221a sEFE>s v* ( 8 \u00b7 ,lm Ut ( 5 6W w 8 \u00b7 cde Ut ( 5 6Wx (9) Using the same methods which were used for force calculation the following momentum coefficients were calculated: ,r T+yz Q ,q ( ,_ V 0 ] w *V,lm U56 9W n{ n| } z (10) ,r * \u221a ,_ w V \u00b7 cde 9 T 4~0 ( + \u0398 \u03bb ] (11) ,r * \u221a ,_ w V \u00b7 ,lm 9 T 4~0 ( + \u0398 \u03bb ] (12) The angle 9 can be seen in Fig.3 and Fig 4. It is important to notice that equations (10) through (12) have two solutions. This is because the rotors spin in different directions as can be seen in Fig. 4. Different rotation directions have the opposite effect on torques which produces the w sign in torque equations. These differences, induced from specific quadrotor construction, along with the augmented momentum equation function (2) provide an improved insight to quadrotor aerodynamics. Regardless of the flying state of the quadrotor, using these equations we can effectively model its behavior" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002814_cdc.2011.6160599-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002814_cdc.2011.6160599-Figure1-1.png", + "caption": "Fig. 1. Illustration of our basic setting.", + "texts": [ + " This implies that when consensus is reached on the utility values, the agents\u2019 states may not necessarily be equal. Additionally, we define a global value that depends directly on all the nodes\u2019 physical states through the so-called global function. By suitable means of communication (or decentralised estimation) either all or just some nodes in the network have access to this global value.1 Finally, we assume that the nodes use the inter-agent communication system to share their current utility value with neighbouring nodes. This set-up is illustrated in Figure 1. The objective is for all agents in the system to reach consensus on their utility values, while also driving a global value to a target value. This should be achieved in a decentralised way, using simple algorithms that will operate in a variety of settings, including time-varying topologies of the communication network, non-linear utility functions that are only known approximately, when not all nodes have access to the global value and when the state updates are not necessarily performed synchronously" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003240_j.jmaa.2012.09.054-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003240_j.jmaa.2012.09.054-Figure3-1.png", + "caption": "Fig. 3. The angle between vector field and normal vectors of boundaries.", + "texts": [ + " An auxiliary circle is also drawnwith center x(t0) and radius \u03b5 (\u03b5 must be small enough such that there is no boundary point in the circle except x(t0)). By moving along the perimeter of the circle from Xb0 to Xj in the clockwise or the counter-clockwise direction, a sequence of regions is observed. The names of these regions are put in an ordered set Xb = Xb0 , Xb1 , . . . , Xbh\u22121 , Xbh from Xb0 to Xbh respectively such that Xbh = Xj. For X\u0304bi X\u0304bi+1 i=1,...,h which surrounded with the auxiliary circle, consider the normal vectors Cbibi+1 . As shown in Fig. 3, since for each X\u0304bi X\u0304bi+1 , we have 0 \u2264 \u03b1 \u2264 180, so the absolute value of the angle between Cbibi+1 and f does not exceed 90\u00b0. This results in Cbibi+1 f \u2265 0. By (9) we have, Vbi+1(x(t0)) \u2264 Vbi(x(t0)), i = 0, . . . , h \u2212 1. So, Vbh(x(t0)) \u2264 Vbi(x(t0)), i = 0, . . . , h \u2212 1 (13) (13) is independent of the direction of moving on the auxiliary circle. So Vj(x(t0)) \u2264 Vi(x(t0)), \u2200i \u2208 Ix, i =\u0338 j. This completes the proof. Lemma 4. If the state trajectory crosses the boundary at x(t0) \u2208 i\u2208Ix X\u0304i, where card Ix(t0) \u227b 2, and lies on the boundary X\u0304j X\u0304j\u2032 , where j, j\u2032 \u2208 Ix(t0) and Fjj\u2032 =\u0338 0, then the inequality (12) holds" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000057_tpas.1980.319632-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000057_tpas.1980.319632-Figure1-1.png", + "caption": "Figure 1. Generalized System for Dynamic Response Analysis", + "texts": [ + " First, the transmission line and the testing procedure will be described to make the test results more meaningful. Description of the Wisconsin Line The tests were performed on a segment of an existing double circuit 138 kV line. The line was originally constructed in 1931 of square based, lattice steel towers on steel grillage foundations. Initially, only one circuit was installed and several years later a second one was added using a different type of conductor. An eight span segment of this line was used for the Wisconsin test, the profile of which is shown in Fig. 1. This segment provided five full spans on one side of the location of the conductor break. It was learned [7,8] that at least three spans were required to not reduce the magnitude of the broken wire loads. 0018-9510/80/0100-0222$00.75(\u00a9)1980 IEEE 223 _ O _5 4C II, @ -N e~~~~ Z n~~~~ Since different conductors were used on the two e0- circuits, the effects of different conductor properties IN o could be studied. Following is a list of the cable L& properties: ,T7 T8 G9 Instrumented Tower Conductor Description Size & Type 397 kcmil ACSR 471A copper/bronze 7-#8 gauge steel Physical Properties Modulus Rated Weight of Elast. Strength Area kgf/m kgf/mm2 kfg m 0.814 1.296 0.479 7100 10500 16000 7393 234.2 6990 143.8 5140 58.1 Fig. 1 Profile of Wisconsin Test Line (Elevations and lengths are given in meters.) The eight spans described in Fig. 1 are essentially all tangent with some small angles. The elevations at the attachment points of the bottom conductor and the span distances between towers are also shown. Towers identified as T3 and T4 were instrumented. Fig. 2 Typical tamgemt tower used im the test lime. Being square based and designed for a single broken wire condition, these towers would, by today's standards, be considered very rigid and strong tangent structures. This fact is further described by the following list of flexibility factors which relate to a longitudinal load applied to the tower arms", + " The test results indicated that in some cases the dynamic loads were amplified by the structure; for example, the measured peak ground-line moment was greater than the ground-line moment corresponding to the peak con- 233 ductor load applied statically. In other cases, the dynamic load was attenuated by the structure. The resonse of the structure to the broken conductor forcing function can be approximated by assuming the structure behaves as an elastic, single degree if freedom system such as that shown in Figure 1. The mass, stiffness, and damping properties of this idealized structure are determined from simplified equations given in Reference 5 and represent the combined properties of the structure and OHGW. The time dependent loading on the structure is the broken conductor forcing function. The ratio of the peak dynamic structure response (e.g., displacement, ground-line moment, member force, etc.) to the static structure response under the peak conductor load will be called the structure response factor (SRF)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001146_1.2991232-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001146_1.2991232-Figure2-1.png", + "caption": "Fig. 2 Schematic of tilting-pad journal bearing with three pads", + "texts": [ + "org/about-asme/terms-of-use a C e d w 3 c s g n a d f t p d t b d b r d l w J Downloaded Fr F\u0304x = Kxx # + i Cxx # X\u03040 + Kxy # + i Cxy # Y\u03040 + j=1 n Kx j + i Cx j \u03040j 27 F\u0304y = Kyx # + i Cyx # X\u03040 + Kyy # + i Cyy # Y\u03040 + j=1 n Ky j + i Cy j \u03040j If assuming that the dimensionless equivalent dynamic stiffness nd damping coefficients of the full bearing are Kkm and km k ,m=x ,y , the dynamic gas-film forces Fx and Fy are also xpressed in the following forms: Fx = Kxx X\u0304 + Kxy Y\u0304 + Cxx X\u0307\u0304 + Cxy Y\u0307\u0304 28 Fy = Kyx X\u0304 + Kyy Y\u0304 + Cyx X\u0307\u0304 + Cyy Y\u0307\u0304 Similarly, the following equations are obtained: F\u0304x = Kxx + i Cxx X\u03040 + Kxy + i Cxy Y\u03040 29 F\u0304y = Kyx + i Cyx X\u03040 + Kyy + i Cyy Y\u03040 From Eqs. 24 , 27 , and 29 , the dimensionless equivalent ynamic stiffness and damping coefficients are expressed as Kkm = Kkm # \u2212 j=1 3 p \u00b7 Kk jK mj \u2212 2Ck jC mj + 2q \u00b7 Kk jC mj + K mjCk j 30 Ckm = Ckm # \u2212 j=1 3 p \u00b7 Kk jC mj + K mjCk j + 2q \u00b7 Ck jC mj \u2212 1 2Kk jK mj here p = K j \u2212 I 2 K j \u2212 I 2 2 + C j 2 31 q = C j K j \u2212 I 2 2 + C j 2 Numerical Results and Discussion A typical hydrodynamic tilting-pad gas bearing under numerial analysis has three pads, as shown in Fig. 2. Its configuration is ame as that of the example in Refs. 14 and 16 . The bearing eometric and operating conditions are listed in Table 1. Its dyamic coefficients are calculated by using the theory described bove. In this study, the changes of the dimensionless equivalent ynamic stiffness and damping coefficients with the perturbation requency are mainly investigated. The results are compared with hose obtained by using the method in Ref. 14 . In this paper, the dynamic coefficients of the bearing and the ads with pad inertia I=0 are given in the form of curves of imensionless stiffness and damping coefficients versus perturbaion frequency" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001054_s12283-010-0034-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001054_s12283-010-0034-3-Figure1-1.png", + "caption": "Fig. 1 The modelled shafts were capable of deflecting about two axes. a Deflection along the Y-axis represents lead/lag motion. b", + "texts": [ + " Before the methodology of this research is explained, a conceptual framework for investigating the mechanisms of clubhead deflection is presented to assist the reader. From a Newtonian perspective, the forces applied to the club by the golfer, along with gravity, cause the shaft to bend during the downswing. A clearer understanding of the source of shaft bending can be gained by resolving the resultant force, applied at the grip end of the club, into a tangential component and a radial component. The tangential component acts in a plane formed by the X and Y axes, while the radial component acts along the Z-axis (Fig. 1). Consider a simplified model of a golf club where a tangential force component acts perpendicularly to the S. J. MacKenzie (&) Department of Human Kinetics, St. Francis Xavier University, 5000, Antigonish, NS B2G 2W5, Canada e-mail: smackenz@stfx.ca E. J. Sprigings College of Kinesiology, University of Saskatchewan, Saskatchewan, Canada shaft at the grip end of the club. Two uniform rigid links are connected by a revolute joint that is spanned by a rotational spring-damper element (Fig. 2). A tangential force is applied by the golfer at the top of Link 1 to accelerate the club laterally, while a stabilizing constraint torque is also applied by the golfer at the top of Link 1 to prevent Link 1 from rotating", + " The model was capable of the four fundamental motions in the downswing: torso rotation, horizontal abduction at the shoulder, external rotation at the shoulder, and ulnar deviation at the wrist. Four muscular torque generators which adhered to the force\u2013velocity and activation rate properties of human muscle were incorporated to add energy to the system. The four segments of the modelled club were connected in series by rotational spring-damper elements (Fig. 4) [5]. The shaft segments were capable of deflecting about two axes (Fig. 1). The model\u2019s goal was to maximize horizontal clubhead speed at impact with the golf ball. An optimization scheme was employed, which used a single activation muscle control strategy where the timing of each muscular torque generator was controlled separately. The optimization search engine was developed by the author and employed an evolutionary algorithm approach, as generally expressed in theory by Michalewicz [6]. Further details on model development, parameters, and optimization can be found in the first paper of this series [1]", + " This would result in the clubhead center of mass no longer being collinear, and radial force would once again exert its influence. Therefore, a second methodology was implemented which allowed both the complete removal and complete isolation of radial force. An optimized simulation of the downswing was generated with the model. The forces and torques applied to the club by the golfer portion of the model were recorded every 10-4 s. The force and torque vectors were each broken down into three components based on the relative reference frame attached to the grip end of the club (Fig. 1). A second confirmatory simulation was conducted with just the four-segment club model, in which the six force and torque measures taken at each time step in the previous simulation served as input. As would be expected, the resulting clubhead speed and clubhead deflection measurements were identical to the first simulation. A third simulation was performed in which the values for the radial force at each time step were set to 0 N, but the other force and torque measures remained the same. This allowed the effect of radial force to be removed from the golf swing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001095_s11044-010-9236-5-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001095_s11044-010-9236-5-Figure5-1.png", + "caption": "Fig. 5 Revolute joint between large deformable and rigid bodies", + "texts": [ + " Furthermore, constraint Jacobian matrices and quadratic velocity vectors specialized for the absolute nodal coordinate formulation need to be developed. Furthermore, it is important to note that the use of such a joint coordinate system allows for cross-sectional deformation at the joint definition point. That is, only nominal orientation of cross section defined using the tangent or cross section frame is constrained. For example, if a large deformable body i formulated using the absolute nodal coordinate formulation is connected to a rigid body j by a revolute joint as shown in Fig. 5, five relative degrees of freedom are kinematically constrained. Since the constraint definition point k of the large deformable body i has 12 degrees of freedom and a rigid body has six degrees of freedom, seven relative degrees of freedom are allowed between the two bodies at the constraint definition point. That is, one degree of freedom is associated with rotation along the axis of revolute joint and the rest of six degrees of freedom are associated with six strain components due to deformation at the constraint definition point", + " On the other hand, in the proposed approach, mapping equations are the only additional constraint equations that need to be added along with the orthonormality conditions, and all the existing joint constraint libraries developed for the rigid body reference coordinates can be directly used without any modifications. This is one of the important advantages of introducing the non-generalized coordinates for joint constraint formulation proposed in this investigation. As discussed in the previous section, if a large deformable body i formulated using the absolute nodal coordinate formulation is connected to a rigid body j by a revolute joint as shown in Fig. 5, only one relative degree of freedom should be allowed between the two bodies at the constraint definition point. The constraint equations given by (27) consist of six equations for CO(eik) = 0, five equations for CJ (pik,qj ) = 0 in the case of revolute joint, and six equations for CR(eik,pik) = 0. Thus, total of 17 equations are defined for the 24 coordinates (12 coordinates for eik , six coordinates for pik , and six coordinates for qj ). Constraining six reference degrees of freedom of one of the bodies, only one relative degree of freedom associated with rotation along the axis of revolute joint is allowed between the two bodies at the constraint definition point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure5-1.png", + "caption": "Fig. 5 Test jig", + "texts": [ + " Table 4 QM test condition Test Item and condition Initial performance Transmission accuracy test Spring rate No-load starting torque Efficiency Vibration test Random vibration: X -, Y -, Z-axis 5\u20132000 Hz, 21 g, 180 s Sinusoidal vibration: X -, Z-axis 10\u2013100 Hz, 25 g, 2 oct/min Thermal vacuum Temperature: test higher side of thermal cycle: +80 + 5/\u22120 \u25e6C, 1 h lower side of thermal cycle: \u221210 + 0/\u22125 \u25e6C, 1 h Vacuum pressure: <10\u22124 Pa Cycle number: 8 cycles Life test Load torque: +/\u221214 Nm (sinusoidal) Input speed: 100 r/min (continuous) Temperature: room temperature (20\u201325 \u25e6C) Vacuum pressure: <10\u22124 Pa Proc. IMechE Vol. 222 Part J: J. Engineering Tribology JET415 \u00a9 IMechE 2008 at IOWA STATE UNIV on October 13, 2014pij.sagepub.comDownloaded from The QM unit was assembled inside a test jig, made for this development only (Fig. 5), throughout the QM test.To check the initial performance, the transmission accuracy, spring rate, no-load starting torque, and efficiency were measured. The test results are presented in Table 5. A vibration test, thermal vacuum test, and life test were carried out at the test facility of JAXA Tsukuba Space Center. The test jig and equipment for the vibration test are shown in Fig. 6. The QM unit was loaded with vibration according to the test conditions exhibited in Table 4. The effect of vibration on the SWG performance was estimated through comparison of the performance data, starting torque, and efficiency before and after the test; no considerable difference between the data before and after testing was found" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003793_j.cja.2013.04.038-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003793_j.cja.2013.04.038-Figure4-1.png", + "caption": "Fig. 4 Topology at reverse connection braking stage.", + "texts": [ + " Its state averaging model is derived as follows: _x \u00bc _x1 _x2 \u00bc 2Rm\u00feRp\u00feRs Lm 0 Kt J Bv J \" # x\u00fe Kex2 2Lm 0 \" # u2 \u00fe DVT\u00fe2DVD 2Lm Td J \" # \u00f07\u00de The braking power is provided with power supply and the buck converter outputs constant capacitor voltage v at the reverse connection braking stage, which must satisfy v> Kex + DVT to eliminate the interphase internal circulation of windings. Besides, the commutation switching transistors VT1 and VT6 execute pulse width modulation in respective conduction region to regulate the braking torque as in Fig. 4 and its instantaneous model can be derived as C dv dt \u00bc i im L di dt \u00bc u4UDC DVT v 2Lm dim dt \u00bc u3\u00f0v\u00fe Kex\u00de 2DVT \u00f02Rm \u00fe Rs\u00deim J dx dt \u00bc Ktim Bvx Td 8>><>>: \u00f08\u00de As averaging is performed over cutting periods, let state variables x1 = v, x2 = i, x3 = im, and x4 = x. Its state averaging model is described as _x \u00bc _x1 _x2 _x3 _x4 26664 37775 \u00bc 0 1 C 1 C 0 1 L 0 0 0 0 0 2Rm\u00feRs 2Lm 0 0 0 Kt J Bv J 266664 377775x\u00fe 0 UDC L 0 0 26664 37775u4 \u00fe 0 0 x1\u00feKex4 2Lm 0 26664 37775u3 \u00fe 0 DVT L DVT Lm Td J 266664 377775 \u00f09\u00de where u3 and u4 are the control inputs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001118_09544054jem1847-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001118_09544054jem1847-Figure3-1.png", + "caption": "Fig. 3 Schematic diagram of rolling contact fatigue test rig", + "texts": [ + " The specimens were cut off and polished for the measurements. Microhardness was measured by the Vickers indenter using a 25 g load. The measurements were taken at the depths of 3, 6, 10, 15, 20, 25, 30, 35, 40, 50, 60, 70, and 80mm. Eight readings were obtained at the same depth from the surface, and the average of those was taken as measurement data. A Zeiss metallographic microscope was used to accurately measure the distance of the indentation from the surface. Rolling contact fatigue tests were performed by using a special test rig (Fig. 3) in a temperature-controlled room set to 25 C. A thrust ball bearing, which has Grade 25 balls of 3.69mm diameter, was inserted between two specimens. The upper specimen was rotated at 1840 rpm, while the lower specimen was fixed in the test rig. Axial loads that produce maximumHertzian stresses of 2724, 3434, 4144, 4854, and 5564MPa were imposed on the upper specimen. The bearing and the specimens were immersed in SAE-30 lubrication oil, which was circulated through a 0.25mm filtered-pump feed system at a rate of 56" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure3-1.png", + "caption": "Figure 3: Principle of Torque Control", + "texts": [ + " Hence the name 'traction drive', which is defined in [2] as: \u201ca power transmission device which utilizes hardened, metallic, rolling bodies for transmission of power through an elastohydrodynamic fluid film\u201d. The force balance in the discs and rollers is shown in figure 2. The application of a castor angle to the roller carriages (as described in Figure 2) enables the rollers to \u2018steer\u2019 to a new angle of inclination and hence Variator ratio. The Torotrak Variator is torque controlled in that the required system torque is set by applying pressure to the pistons connected to the rollers and the Variator follows the ratio automatically [3]. Figure 3 explains this approach using a simplified single roller model. Applying a reaction force F to the roller causes a reaction torque (Ta and Tb) at the Variator discs and consequently an acceleration of the two inertias (engine side inertia A and vehicle side inertia B). This may change the speed of the engine and/or vehicle inertia resulting in a change of Variator ratio. Due to application of the castor angle, this ratio change happens automatically. In the Torotrak Variator design described above, reaction force is applied hydraulically to individual roller carriage pistons" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002727_1.4023084-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002727_1.4023084-Figure3-1.png", + "caption": "Fig. 3 Free vibrations of a one-degree-of-freedom system decay exponentially", + "texts": [ + " (5) can be written as \u20ac h0 s\u00f0 \u00de=X2 m \u00fe C _ h0 s\u00f0 \u00de \u00fe K h0 s\u00f0 \u00de h0S\u00bd \u00bc 0 (7) 021501-2 / Vol. 135, APRIL 2013 Transactions of the ASME Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 07/28/2013 Terms of Use: http://asme.org/terms Eq. (7) is rewritten in terms of xn and n as \u20ac h0 s\u00f0 \u00de \u00fe 2nxn _ h0 s\u00f0 \u00de \u00fe x2 n D h0 s\u00f0 \u00de \u00bc 0 (8) where D h0 s\u00f0 \u00de \u00bc h0 s\u00f0 \u00de h0S;xn \u00bc ffiffiffiffiffiffiffiffiffiffi KX2 m q ; n \u00bc CX2 m=\u00f02xn\u00de, representing the displacement of mutual approach, natural frequency and damping ratio of free vibration, respectively. As shown in Fig. 3, the amplitudes of D h0 decay in the proportion of e nxns. Thus, the logarithmic decrement d can be obtained as d \u00bc ln A1=A2\u00f0 \u00de \u00bc nxnsn \u00bc 2pn= ffiffiffiffiffiffiffiffiffiffiffiffiffi 1 n2 q (9) where A1 and A2 are the amplitudes of D h0 and sn denotes the period of oscillation. As the model in this paper is linearized for a nonlinear system, the response curve is not in agreement with the exponentially decay curve well. Thus, the average value of the equivalent logarithmic decrement d in two periods can be more reasonable than its value in one period" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000540_insi.2008.50.4.195-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000540_insi.2008.50.4.195-Figure2-1.png", + "caption": "Figure 2. Two-stage gearing system with six torsional degrees of freedom, electric motor moment Ms(\u03d51) and external load moment Mr; system consists of: rotor inertia Is, gear inertia of first stage I1p, I2p, gear inertia of second stage I3p, I4p, driven machine inertia Im, gearing stiffness kz1, kz2 and damping Cz1, Cz12, gearing stiffness forces F1, F2, and damping forces F1t, F2t, internal moments in first shaft M1; M1t (M1t coupling damping moment), shaft stiffness k1, k2, k3, internal moment in second and third shaft M2 and M3", + "texts": [ + " Mathematical modelling and computer simulation for revealing vibration signal properties To increase the concessions of the relation between factors having influence on the vibration signal and signal form, mathematical modelling and computer simulation (MM and CS) is used. A detailed description of the mentioned relation is presented in the literature[1 to 9]. In investigating different properties of vibration signals by MM and CS, different models of gearboxes are taken into consideration. The models can be roughly divided into ones where only torsional vibration is taken into consideration and models where torsional together with lateral vibration is taken into consideration. Figure 2 shows a model where only the torsional vibration is considered. To use the model given in Figure 2 for computer simulation, it has to be written as the equation of motion[2]. In the equations of motion, physical values described in Figure 2, which express the design and operation factors, are included. The design and operation factors are given by the electric motor moment Ms(\u03d51) and the external load moment Mr. The system consists of: the rotor inertia Is, the gear inertia of the first stage I1p, I2p, the gear inertia of the second stage I3p, I4p, the driven machine inertia Im, the gearing stiffness kz1, kz2 and damping Cz1, Cz12, the gearing stiffness internal forces F1, F2, and damping internal forces F1t, F2t, the internal moments in the first shaft M1; M1t (M1t - the coupling damping moment), the shaft stiffness k1, k2, k3, the internal moment in second and third shaft M2 and M3", + " Vibration signal properties, methods of signal analysis, degradation scenarios and prognosis, diagnostic inferring process Taking into consideration the vibration signal properties, the means of signal analysis, the degradation scenario and prognosis, and the diagnostic inference process, a diagnostic method for the condition monitoring can be developed. The general aim of the gearbox condition monitoring is the assessment of condition change using the vibration signal. There are different starting assumptions taken for the condition monitoring. They can be described as is given in this paper in the chapter about factor analysis. As will be shown, usually, only some factors presented in Figure 1 and Figure 2 are taken into consideration. Referring to the literature[4][21] it can be seen that the aim of the presented diagnostic method is the diagnostic assessment of the gearbox condition described by the limited imperfections given by the dimension and shape deviations (according to Polish Standards (PS) similar to ISO). Figure 3 presents the vectors of the gear condition with the modified, unmodified and pitted gears: W, 1, 2 (class 6 PS) \u2013 modified gears; 3, 4, 5 \u2013 unmodified gears (class 8, PS); 6, 7, 8 \u2013 unmodified gears (class 7, PS); 9 \u2013 unmodified gears (class 6, PS); 10-pitted gear" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000325_bm060877b-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000325_bm060877b-Figure2-1.png", + "caption": "Figure 2. Schematic of the beam passing through a fiber gauge volume where d is the fiber diameter and Ag is a two-dimensional projection of gauge volume.", + "texts": [ + " To calculate gauge volume weighted averages, it is necessary to determine the amount of material at each beam position across the fiber. Gauge volume is the amount of material through which the X-ray beam passes. By assuming that the fiber is cylindrical and that the beam spot size is constant across the fiber width, the calculation of gauge volume can be performed using the equation where and Ag is the gauge volume, d is the fiber diameter, and x is the position of the fiber in the beam, as shown in Figure 2. Therefore, the weighted average is given by the equation Ag ) 2 \u222bx1 x2 x(d/2)2 - x2 dx (2) -d/2 < x1 < x2 < d/2 \u2329sin2 \u03b8\u232aaverage ) \u2211 n (\u2329sin2 \u03b8\u232anAg,n) \u2211 n Ag,n (3) \u2329sin2 \u03b8\u232a ) \u222b0 \u03c0/2 F(\u03b8) sin3 \u03b8 d\u03b8 \u222b0 \u03c0/2 F(\u03b8) sin \u03b8 d\u03b8 (1) where \u2329sin2 \u03b8\u232aaverage is the calculated average orientation parameter, \u2329sin2 \u03b8\u232an is the orientation parameter at the nth position on the fiber, and Ag,n is the nth gauge area for that position. Determination of Fiber Orientation during Deformation Using X-ray Diffraction. WAXD deformation studies of all samples were carried out on the protein crystallography beamline (station 14" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001143_j.commatsci.2009.11.013-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001143_j.commatsci.2009.11.013-Figure2-1.png", + "caption": "Fig. 2. Schematic of test plate.", + "texts": [ + " The solution domain (300 mm 200 mm 9 mm) is discretised using eight-node hexahedral elements. As shown in Fig. 1, finer mesh is used in the weld zone and its surroundings, and a coarser mesh is used for the rest of the structure, so that the contradiction between the computational precision and the computational time can be solved appropriately. dual stress distribution for test case A2 (GMAW-P). 1000 s after welding (10 ). The workpiece material is mild steel Q235 with dimensions of 300 mm in length, 200 mm in width and 9 mm in depth. As illustrated in Fig. 2, bead-on-plate welding was performed at the center of the specimen. During welding process two ends of the plate were constrained while they were taken off after welding. Both GMAW-P and Laser + GMAW-P hybrid welding were conducted on such test plates. To compare the residual stresses caused by GMAW-P and Laser + GMAW-P hybrid welding processes, the heat input should be so selected that the weld penetration is nearly same for both processes. By adjusting welding current or welding speed, approximately same weld penetrations were obtained for GMAW-P and Laser + GMAW-P hybrid welding processes under two groups of experiments with following process parameters", + " seen that hybrid weld is more laser-like at the bottom and more GMAW-like on the top due to the two process combinations. Table 3 lists the predicted and measured weld dimensions for all four welding conditions. The computed and experimental measured weld geometry and size are in agreement with each other. Thus, the thermal analysis results can be used for thermomechanical analysis. Then, the solutions of nodal temperature distributions are used as input to thermomechanical stress analysis. As shown in Fig. 2, the residual stress distributions along two lines on the workpiece, LINE_X (perpendicular to the weld line) and LINE_Y (parallel to the weld line), are computed. Figs. 9 and 10 compare the calculated and experimental measured residual stress distributions on LINE_X for test cases A1 (GMAW-P) and B1 (Laser + GMAW-P), respectively. The computed residual stresses agree with the experimental date, although the error of the transverse stresses is a little larger. The transverse stress consists of two parts, one part is from the longitudinal shrinkage in the weld and its surrounding plastic deformation zone, and another part is induced by the non-synchronous transverse shrinkage in the weld and its surrounding plastic deformation zone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001472_1869983.1870008-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001472_1869983.1870008-Figure2-1.png", + "caption": "Figure 2. Depth adjustment system and AQUANODE.", + "texts": [ + " We have developed an inexpensive underwater sensor network system that incorporates the ability to dynamically adjust its depth. The base sensor node hardware is called the AQUANODE platform and is described in detail in [13, 15, 35]. We have extended this basic underwater sensor network with autonomous depth adjustment ability and created a five node sensor network system, whose nodes move up and down in the water column under their own control. Each node costs under two thousand dollars. Here we will briefly summarize the system and describe some details of the winch-based depth adjustment system. Figure 2 shows a picture of two AQUANODES with the depth adjustment hardware. The AQUANODE operating system runs on an NXP LPC2148 ARM7TDMI processor clocked at 60MHz. This processor has 40kB of ram and 512kB of on-chip flash. A SD card slot allows logging of gigabytes of data onboard. Each node has a pressure and temperature sensor as well as inputs for both analog and digital sensors connected via an underwater connect. Examples of sensors we have connected include CDOM, salinity, dissolved oxygen, and cameras" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001814_tmag.2010.2043075-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001814_tmag.2010.2043075-Figure2-1.png", + "caption": "Fig. 2. Simplified intersection of master and slave domain and integration points on and on .", + "texts": [ + " The continuous shape functions are of third polynomial order. Their numerical integration, compared to the discontinuous shape functions of first order, require a slightly higher computational effort. Anyhow, numerical experiments have not revealed any disadvantages in using the discontinuous basis functions. IV. IMPLEMENTATION ASPECTS The described approach has been implemented within the FEM-package MOOSE [10]. In rotational motion problems, discretization inevitably polygonizes the interfaces and , Fig. 2. The simple rotation mapping (23) is, therefore, in general, not a mapping as it should be, i.e., the image of the integration Gauss point is not found on . A straightforward solution is to define as the orthogonal projection of on . This minimizes the distance between and . The required mapping is thus the composition of this projection with . The shape function can then be evaluated at . Since the integrand in (16) is no longer -continuity, it is advisable to use more integration points than for -continuity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000176_09544054jem944-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000176_09544054jem944-Figure1-1.png", + "caption": "Fig. 1 Experimental set-up and parameter definitions for CO2 Laser\u2013GMA hybrid welding", + "texts": [ + "comDownloaded from with a focal length of 250mm; diameter of the laser spot was approximately 0.4mm at the focal point. A Fronius TPS 5000, which is a constant voltage power source with a rated output of 500 A, was used for GMAW. To secure a stable wire-feeding condition, a water-cooled push\u2013pull torch was used. All trials were conducted with d.c., electrode-positive and constant voltage configuration. The experimental set-up, including a self-designed hybrid welding head, and the definitions of welding parameters are shown in Fig. 1. A number of trials were conducted to optimize and determine the process parameters of hybrid welding. Welding parameters and their values used in the experiments are listed in Table 2. Welding speed was fixed at 1.5m/min. At present, 8mm-thick A-grade steel is usually welded with a speed of 0.75m/min by submerged arc welding (SAW) in the shipbuilding industry [16]. From an economic point of view, to substitute hybrid welding for the conventional SAW process, this research adopted a 1.5m/min welding speed to provide productivity four times greater than that of SAW" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002572_am2004637-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002572_am2004637-Figure1-1.png", + "caption": "Figure 1. Schematic diagram of two porous sol gel glass layers coated on a planar ITO electrode. PANI nanowires are electrochemically polymerized from the ITO surface through the pores of the lower sol gel layer. A planar supported lipid bilayer (PSLB) is deposited on the upper sol gel layer. Valinomycin, an ionophore, and CCCP, a protonophore, are dissolved in the PSLB, where P , HP, and VAL represent deprotonated CCCP, protonated CCCP, and valinomycin, respectively. A transmembrane pH gradient drives transmembrane proton transport via HP/P and aided by VAL. This process decreases the pH below the PSLB, which is detected as a change in the potential of the PANI/sol gel/ITO electrode.", + "texts": [ + "39,40 PANI can be synthesized within sol gels by first allowing the aniline monomers to diffuse into the porous glass network, followed by electrochemical polymerization,41,42 or by premixing the monomers with the sol gel precursor solution, followed by deposition onto an electrode and subsequent electrochemical polymerization.36,43 The latter strategy was employed here to prepare a sol gel layer on ITO functionalized with PANI nanoelectrodes. A second thinner sol gel layer was then applied to create a smooth hydrophilic surface suitable for deposition of a planar supported lipid bilayer (PSLB) by vesicle fusion. The overall structure of the supramolecular assembly, which incorporates ionophores into the membrane, is shown in Figure 1. The feasibility of using this PANI nanoelectrode/sol gel hybrid layer to monitor ionophore-facilitated proton transport across the PSLB was demonstrated. \u2019EXPERIMENTAL SECTION Reagents. The following chemicals were purchased from Aldrich and used without further purification: tetraethyl orthosilicate (TEOS), 99.999%;methyltriethoxysilane (MTES), 99%; valinomycin, 98.0%; and carbonyl cyanide 3-chlorophenylhydrazone (CCCP), 98.0%. Aniline (Aldrich, 99.5%) was distilled before use. Deionized (DI) water was obtained from a Barnstead Nanopure system with a measured resistivity of 18", + " Transmembrane proton transport aided by lipophilic ionophores was examined to further characterize the pH sensing properties of these multilayer assemblies. Valinomycin, a K+ ionophore, and CCCP, a protonophore, were inserted into PSLBs on CL/ PANI/sol gel/ITO electrodes to measure \u0394pH-driven proton transport. Ionophore-mediated transmembrane proton transport using valinomycin and CCCP has been studied in liposomes,61,62 and the proton transport mechanism has been described by Ahmed and Krishnamoorthy.62 64 The proposed mechanism, illustrated in Figure 1, involves cotransport of K+ in the opposite direction to counter the charge imbalance generated by proton transport. To our knowledge, valinomycin/CCCP-mediated proton transport across lipid bilayers has only been studied using liposomes. The behavior of each component of the molecular assembly (Figure 1) was tested individually during each experiment to ensure the proper functioning of the entire assembly, as follows: (1) The pH response of the CL/PANI/sol gel/ITO electrode was tested before PSLB deposition. Typically, a near Nernstian response with a slope of 38 mV/pH unit was observed (Figure 6b). (2) The barrier properties of the PSLB were tested as described above. Typically, the variation in cell potential was less than 2 mV when the pH of the buffer above the PSLB was changed, as shown in section (a) of Figure 7", + " The time-dependent potentiometric response curve was converted into a temporal pH response curve using the calibration curve obtained for electrodes lacking PSLBs (Figure 6b). In panel (b) of Figure 8, the\u0394pH relaxation data are plotted for the buffer change from pH 7 to 6.5. A similar curve was observed for the pH change from 6.5 to 6 (data not shown). 2684 dx.doi.org/10.1021/am2004637 |ACS Appl. Mater. Interfaces 2011, 3, 2677\u20132685 Asmall pHgradient across a lipidbilayer shoulddecay according to58 \u0394pH\u00f0t\u00de \u00bc \u0394pH\u00f0t \u00bc 0\u00de exp\u00f0 t=\u03c4\u00de \u00f04\u00de where t is the time, \u03c4 is the lifetime, and\u0394pH is thepHdifference across the bilayer. For the planar assembly shown in Figure 1,\u0394pH = pHu pHa where the subscripts u and a refer to below and above the PSLB, respectively. The change in pH below the PSLB per unit time can be expressed as58 dpHu dt \u00bc Jnet A VB \u00f05\u00de whereV is the aqueous volume below the PSLB, B = (d[H+]/dpH) is the buffer capacity in that volume, and A is the PSLB surface area; eqs 3 5 allow calculation of Jnet and Pnet from the experimental data. The data in panel (b) of Figure 8 were fit to a single exponential decay from which \u03c4 = 270 s and (dpHu)/(dt) = 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001684_wcm.873-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001684_wcm.873-Figure2-1.png", + "caption": "Figure 2. An inverted pendulum system.", + "texts": [ + " Equation(16) (18) It is worth noting that Theorems 3 and 4 are one of several possible ways to formulate the H2 output tracking control problem as iterative convex programming problems based on Theorem 2. The H2 output tracking control problem can also be formulated as bilinear matrix inequalities (BMIs) based on Theorem 2 and then be solved via LMI-based algorithm, e.g., V-K iteration algorithm [12]. Simulation examples are given in this section to support the developed theory and compare with the classicalH2 control method. Consider an inverted pendulum system shown in Figure 2, where \u03b8 is the angular position of the pendulum, and u is the input torque. The state variables are chosen as [\u03b8T \u03b8\u0307T]T. The output is y = [\u03b8]. The parameters here are: m = 0.01 kg, L = 0.1 m. The state feedback controller is designed using the discrete-time model. Hence, the 1112 Wirel. Commun. Mob. Comput. 2011; 11:1107\u20131116 \u00a9 2009 John Wiley & Sons, Ltd. DOI: 10.1002/wcm linearized discrete-time system (sampling time Ts = 0.01 s) is x(k + 1) = Adx(k) + Bdu(k), y(k) = Cdx(k) where Ad = [ 1.0049 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002066_j.triboint.2012.05.014-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002066_j.triboint.2012.05.014-Figure3-1.png", + "caption": "Fig. 3. (a) Schematic drawing of the tribological contact pairing a textured pin with geometry for the 3D-CFD computation, (b) the computational domain with the boun domain with the boundary conditions for the 2D two-phase flow simulations.", + "texts": [ + " The channels were oriented with an angle of 451 to the sliding direction. Supporting FE simulations were conducted using the CFD tool COMSOL 4.1 for single and multiphase flow modeling. Assuming an incompressible fluid of density r and dynamic viscosity Z and neglecting body forces, the Navier Stokes equations (Eq. (1)) describe the fluid motion based on the velocity field u and the pressure p. r @u @t \u00fe\u00f0u r\u00deu \u00bc pI\u00feZ\u00f0ru\u00feur\u00de \u00f01\u00de r u\u00bc 0 \u00f02\u00de A single unit cell of the texture pattern in the middle of the tribological contact served as the model geometry (Fig. 3a). It contains the microchannels and the oil film between the pin and the disc. The disc itself was represented by a moving wall boundary condition (Fig. 3b) on the lower side of the lubricant film. On all channel walls \u2018\u2018no-slip\u2019\u2019 boundary conditions were applied. On opposing sides periodic flow conditions were used (Fig. 3b). Simulation parameters are given in Table 1. Only the case of film thickness h\u00bc 0 mm was investigated here to study the oil flow and the pressure distribution in the microchannels when a continuous lubrication film is lost. Motivated by experimental observations of gas bubbles inside the channels, additionally, two phase flow simulations in two dimensions were conducted mimicking only the flow within the channels. They were based on the implementation of [22] and used only to trace the shape of the second phase. Fig. 3c shows the model geometry with the boundary conditions. Here, the flow was initiated by an inflow and outflow condition at the upper and lower end of the domain. The sides were supplied with periodic flow conditions. A wetting wall condition as provided in the software tool was used for the contact between the wall and the second phase. A wetting angle of 111 was used, which was taken from wetting experiments with oil FVA No. 3 on a polished steel surface [13]. Table 1 Parameters of the 3D-CFD simulations", + " Only for the 300 mm wide channels the bubbly region was reduced to a small size and to partial coverage of the channels just like for the thinner channels shown in Fig. 4b, d, but the bubbly region could not be completely eliminated. a polished disc designating the basis cell of the texture which served as model dary conditions for the 3D single-phase flow simulations, (c) the computational The 3D Navier\u2013Stokes equation has been solved for an incompressible fluid within the channel geometry shown in Fig. 3b and the properties of FVA oil No. 3 shown in Table 1. Fig. 6a gives the pressure distribution from the inlet (I) to the outlet (O) from point A to point C (Fig. 3b) at the ceiling of the channel. The distance has been normalized by the length of the plateau edge. Different channel geometries similar to the experimental ones are shown in Fig. 6a. The pressure is symmetric to the point B, with higher values on the upstream end (pointing towards A) and lower values on the downstream end (pointing towards C). The highest pressure difference between the front and back end of the plateau was observed in the 300 mm wide channels. There, the pressure changed the sign from positive (compression) on the upstream end of the plateau to negative (tension) on the downstreeam end" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003077_j.mechmachtheory.2013.09.012-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003077_j.mechmachtheory.2013.09.012-Figure1-1.png", + "caption": "Fig. 1. Experimental test rig.", + "texts": [ + " In this contribution, each node is considered as a possible balancing plane. In this paper, the method proposed and the system studied are presented first; then, the results obtained for the model identification are shown; for this step, only experimental investigations are reported in this work. Once the rotor is identified, numerical and experimental investigations are performed for the optimization of the unbalanced distribution. Finally, the results obtained are discussed in the conclusion. 2. System configuration The system studied (Fig. 1) is a test machine composed of a horizontal flexible shaft of 0.04 m diameter containing two rigid discs. The rotor is driven by an electrical motor that can accelerate the shaft until a rotation of 10,000 rpm. The shaft is supported by bearings located at its ends, as follows: a roller bearing (B2) at one end and two ball bearings at the other end (B1). The roller bearing is located in a squirrel cage attached to the framework of the test bench by three identical flexible steel beams. The Electro-Magnetic Actuator (EMA) located on the external cage constitutes a smart active bearing and provides nonlinearity in the dynamics of the system. The active bearing is placed close the drive end. The displacements are measured by using four proximity sensors (Vibrometer TQ 103) arranged perpendicularly in two measurement planes located along the y-axis, namely, measurement planes #1 and # 2 (Fig. 1). The geometries of the actuators are summarized in Fig. 2. Since an EMA can only produce attractive forces, four \u201cidentical\u201d EMA supplied by constant currents are utilized. Each EMA is composed of a ferromagnetic circuit and an electrical circuit. The ferromagnetic circuit has two parts: an (E) shape, which receives the induction coil, and an (I) shape, which is fixed to the squirrel cage. Both parts are made of sets of insulated ferromagnetic sheets. The quality of the ferromagnetic circuit alloy is considered high enough and the nominal air gap between the stator and the beam is small enough to consider magnetic loss as negligible", + " The system (without the hybrid bearing) is linear; the cross stiffness terms due to the localized non-linearity were taken into account in the second member of the equation of motion and are introduced as restoring force. The objective function is given by Eq. (3). It takes into account the norm of the difference between the experimental FRFs and those generated in the vicinity of the first two vibrating modes for each direction. OF \u00bc X2 j\u00bc1 X2 i\u00bc1 norm FRFmeas i \u03c9a\u22ef\u03c9b\u00f0 \u00de\u2212FRFmodel i \u03c9a\u22ef\u03c9b\u00f0 \u00de norm FRFmeas i \u03c9a\u22ef\u03c9b\u00f0 \u00de 0 @ 1 A j ; \u00f03\u00de j represents the two sensors along each direction (the planes 1 and 2 as shown in Fig. 1, and nodes #22 and #36 as shown where in Fig. 4, respectively); i stands for the considered mode (twomodes are taken into account for each direction);\u03c9a and\u03c9b indicate the frequency range used. 4.2. Identification of the unbalance distribution The unbalancing identification process was done in two steps. In the first step, the shaft was divided into several intervals. It was considered that 2 unbalances per interval could exist. The first step was not used to determine the exact correction weights; it was dedicated to providing a representation of the unbalance distribution and the choice of the best intervals that should be used to apply the correction weights" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure11.2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure11.2-1.png", + "caption": "Fig. 11.2 Robot SMART 3-S (VRML model)", + "texts": [ + " Section 4 gives details on the overall software architecture and Section 5 illustrates the remote operator teleoperation interface. The architecture of the robotic teleoperation system (shown in Fig. 11.1) consists of the following components: \u2022 An industrial robot COMAU SMART 3-S \u2022 A robot controller COMAU C3G-9000 \u2022 A PC operating as a server, connected to the robot controller via a serial RS-232 link and to the Internet \u2022 Three web cams with dedicated communications 252 G. Ferretti et al. BookID 175907_ChapID 11_Proof# 1 - 15/4/2009 BookID 175907_ChapID 11_Proof# 1 - 15/4/2009 The SMART 3-S (Fig. 11.2) is an industrial robot with six rotational degrees of freedom, mainly used for arc welding, manipulation, painting and water jet processing. It carries a payload up to 6 kg and is characterized by a repeatability of 0.1 mm. The end effector is a pneumatic gripper, made up by two opposing parallel fingers. The gripper is controlled through an electric valve and is endowed with two proximity sensors, in order to monitor the states of the grasp operation: open gripper, closed gripper and grasped object, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003102_tcst.2012.2216268-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003102_tcst.2012.2216268-Figure2-1.png", + "caption": "Fig. 2. Two-link experimental apparatus, ideal frequency response, and realistic frequency response. (a) Two-link flexible manipulator apparatus. (b) Frequency response of the ideal two-link system and the two-link system that uses F(s) to estimate rates.", + "texts": [ + " Parameterizing in terms of Ko alone is indeed restrictive, as is simply mimicking an unconstrained H2. However, we do so in order to pose a tractable convex optimization problem constrained by LMIs. The FF SPR/BR controller synthesis method of Section III will be used to control a two-link flexible manipulator. This system is nominally passive; however, as we will show, a simple filter dynamics destroys the passive IO properties of the system, rendering it hybrid passive/finite-gain in nature. Consider the two-link flexible manipulator in Fig. 2(a). The first link is 210.00-mm long, 1.27-mm thick, and 76.20-mm high. The second link is 210.00-mm long, 0.89-mm thick, and 38.1-mm high. Each link is made of steel and has a strain gauge at its base. The manipulator is manufactured by Quanser Consulting Inc. Additional information can be found in [36]. The dynamics of the system is described by the following second-order nonlinear matrix differential equation [5], [6]: M(q)q\u0308 + Dq\u0307 + Kq = B\u0302\u03c4 + fn(q, q\u0307) (20) where M = MT > 0 is the mass matrix, D = DT \u2265 0 is the damping matrix, K = KT \u2265 0 is the stiffness matrix, and B\u0302 = [1 0]T. The column matrix q = [\u03b8T qT e ]T is composed of the joint angles \u03b8 = [\u03b81 \u03b82]T and the elastic coordinates associated with the discretization of each link, qe. The term fn stems from nonlinear inertial forces. Joint torques \u03c4 = [\u03c41 \u03c42]T are applied by motors at the base of each link. The apparatus in Fig. 2(a) is equipped with two encoders, one affixed to each joint [36]. The joint encoders provide \u03b81 and \u03b82. As such, proportional control can be implemented easily. For rate control to be implemented, \u03b8\u0307 must be made available. Unfortunately, our apparatus is not equipped with any sort of rate sensor, and as such \u03b8\u0307 will be acquired through some sort of differentiation. If perfect differentiation were possible, then the mapping \u03c4 \u2192 y where y(s) = s\u03b8(s) would be passive. In practice, perfect differentiation is not possible and can only be approximated", + " Note that we are deliberately writing the linearized motion equations in this form because numerical computations tend to be much more stable (as we found out when actually computing controllers using the method in Section III). Given the above state-space form, the relation between x, \u03b8 , and \u03b8\u0307 is \u03b8 = [ B\u0302TQe \u22121 0 ] \ufe38 \ufe37\ufe37 \ufe38 Cp x, \u03b8\u0307 = [ 0 B\u0302TQe ] \ufe38 \ufe37\ufe37 \ufe38 C x. (23) Let \u03b8(s) = Gp(s)\u03c4 (s) = Cp(s1 \u2212 A)\u22121B\u03c4 (s), and \u03b8\u0307(s) = G(s)\u03c4 (s) = C(s1 \u2212 A)\u22121B\u03c4 (s), where G(s) := sGp(s). The frequency response of the ideal (linearized) system G(s) is shown in Fig. 2(b). The linearization is performed about \u03b8d = [\u2212\u03c0/4 0]T. Within Fig. 2(b) is plotted the maximum singular value of G(s) and the minimum Hermitian part as a function of frequency. The maximum singular value of G( j\u03c9) is \u03c3\u0304 (G( j\u03c9)) = \u221a \u03bb [ GH( j\u03c9)G( j\u03c9) ] , while the minimum Hermitian part is (1/2)\u03bb [ G( j\u03c9) + GH( j\u03c9) ] . Clearly, the linearized system is PR over all frequencies owing to the fact the Hermitian part is positive over all frequencies. This result is expected. Now consider the IO mapping where \u03b8\u0307 is not directly measured, but acquired via differentiation using F(s), i.e., y(s) = G1(s)\u03c4 (s) where G1(s) = F(s)Gp(s). The frequency response of this transfer matrix is also plotted in Fig. 2(b). The system G1(s) is PR over a specific frequency range; below approximately 100 rad/s the transfer matrix has a Hermitian part that is positive, and hence PR. Above 100 rad/s, the system is no longer PR (i.e., the Hermitian part is negative) but is BR. The system is clearly hybrid passive/finite-gain possessing a frequency response that is FF PR/BR. Assuming that our linearized model accurately approximates the nonlinear system, by using the hybrid passivity/finite-gain stability theorem this system can be stabilized by an FF SPR/BR controller", + " The two controllers within the scheduling algorithm, G21(s) and G22(s), will each be designed about a specific linearization point: G21(s) about set point 1, and G22(s) about set point 2. Set point 1 corresponds to [\u2212\u03c0/4 0]T rad, while set point 2 corresponds to [(\u03c0/4) (\u03c0/3)]T rad. The weights used for controller synthesis are B1 = 10 [ B 0 ] C1 = \u23a1 \u23a3 100Cp 2C 0 \u23a4 \u23a6 D12 = [ 0 1 ] D21 = [ 0 1 ] (24) where B, Cp , and C are given in (22) and (23). To design the FF SPR/BR controller, both \u03c9c and \u03b31 must be estimated. From Fig. 2(b), \u03c9c = 100 rad/s while \u03b31 = 1.25 rad/(N \u00b7 m \u00b7 s). Note that these values are estimated based on a linearization and an assumed filter F(s); the true nonlinear high-frequency gain may not be the \u03b31 we have chosen. However, with no way to calculate a true nonlinear gain, we resort to estimating the high-frequency gain in this way. The frequency response of the FF SPR/BR controllers synthesized about set points 1 and 2 using the scheme presented in Section III are shown in Fig. 4(a) and (b). The frequency responses of the H2 controllers used as the basis controllers for the the FF SPR/BR controllers are also shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001095_s11044-010-9236-5-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001095_s11044-010-9236-5-Figure3-1.png", + "caption": "Fig. 3 Joint constraint using joint coordinate system", + "texts": [ + "1 Joint coordinate systems Most of joint constraint equations can be defined by a combination of several basic constraints given, for example, as [1, 2]: CS = ri \u2212 rj \u2212 c = 0, CN1 = vi \u00b7 vj = 0, CN2 = vi \u00b7 rij = 0 (11) where c is a constant vector, vi and vj are vectors defined on bodies i and j used to describe the orientation of the joint axis, and rij = ri \u2212 rj . The preceding constraint equations are, respectively, used to describe relative displacement and orthogonal conditions of vectors on two bodies. In order to impose such constraint equations between a large deformable body formulated using the absolute nodal coordinate formulation and a rigid body, joint coordinate systems are defined at the constraint definition point as shown in Fig. 3, and connectivity conditions are described with respect to these coordinate systems [5]. When body i is defined using the absolute nodal coordinate formulation and a joint is defined at node k, the vector vi in (11) is defined as vi = Aik v\u0304ik with constant vector v\u0304ik defined with respect to the joint coordinate system. Since the orientation and deformation of element i are parameterized using the position vector gradient coordinates as shown in (2), the orthonormal orientation matrix can be, in principle, obtained using the Polar Decomposition Theorem as Jik = AikUik (12) where Jik is a matrix of position vector gradient at node k, Aik is an orthonormal orientation matrix, and Uik is a symmetric stretch tensor", + " 4(b), X-axis of the tangent frame, which is defined by a unit vector iikt in (17), is always tangential to the beam centerline, and this vector is given as iikt = \u2202rik \u2202sik = \u2202rik \u2202xi ( \u2202sik \u2202xi )\u22121 = \u2202rik/\u2202xi |\u2202rik/\u2202xi | (18) where sik represents the arc-length coordinate along the beam centerline at the deformed configuration and is written as dsik = \u221a( \u2202rik \u2202xi )T ( \u2202rik \u2202xi ) dxi = \u2223\u2223 \u2223\u2223 \u2202rik \u2202xi \u2223\u2223 \u2223\u2223dxi. (19) If deformation of a cross section is small, the tangent frame can be used to describe the orientation of the cross section at the constraint definition point. 3.2 Constraint formulations When a flexible body i modeled using the absolute nodal coordinate formulation is connected to a rigid body j at point P by a revolute joint as shown in Fig. 3, the constraint equations can be defined as C ( eik,qj ) = \u23a1 \u23a3 rik P \u2212 rj P vik 1 \u00b7 vj vik 2 \u00b7 vj \u23a4 \u23a6 = 0 (20) where eik is a vector of nodal coordinates at point P defined on node k of element i, and qj is a vector of the generalized reference coordinates of a rigid body j . The vector rik P defines the global position at P ik on flexible body i, while rj P defines the global position at P j on a rigid body j ; vik 1 and vik 2 are two vectors orthogonal to an orientation vector of joint axis vik defined on the flexible body, and vj defines the orientation of the joint axis on a rigid body j " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure2-1.png", + "caption": "Fig. 2. Comparison of several types of in-pipe mobile robots.", + "texts": [ + " It is inspired by the motion of spirochetes. Many different microorganisms move as shown in Fig. 1(a); the body and a flagellum for propulsion are separate. These microorganisms use only the flagellum for propulsion. In contrast, a spirochete has a helical body and rotates about the axis of its body to generate propulsive power, as shown in Fig. 1(b). To generate rotational torque of the entire body spirochetes use helical rotating motion. This motion should be highly efficient because the entire body generates the propulsive force. Figure 2 compares a helical rotation in-pipe mobile robot and other types of in-pipe mobile robots. A wheeled robot[1][3] can move quickly because wheels provide highly efficient propulsion. However, friction is generated if its body touches 978-1-4244-7709-8/10/$26.00 \u00a92010 IEEE 313 the inner wall of a pipe. To avoid this friction, wheels must be placed on every side of the robot, complicating the body mechanism. Moreover, the unlimited rotary mechanism is difficult to seal off from hazardous environments" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000083_j.apm.2007.07.007-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000083_j.apm.2007.07.007-Figure1-1.png", + "caption": "Fig. 1. Gear contact geometry definitions for the derivation of the meshing equations.", + "texts": [ + " Given that the module is a characteristic constant in any gear pair that both gears share, it makes sense to non-dimensionalise the geometrical analysis, so that all lengths are expressed as non-dimensional multiples of the gear module. Non-dimensional gear mesh analysis results can therefore be used to describe a family of geometrically similar gear pairs, instead of just a single gear pair, lending an insightful generality to the calculations. Let us assume a pair of gears 1, 2 and their respective teeth (correspondingly marked by indices 1, 2) in contact as shown in Fig. 1. Assuming a non-dimensional Cartesian coordinate system the gear profiles are considered to lie on the x1x2-plane and their axes of revolution are parallel to x3-axis. The orientation of axes x1 and x2 as well as the coordinate system origin may be chosen arbitrarily. If gear 1 is regarded as the reference gear and conjugate gear action is assumed, then tooth 2 should occupy a nominal angular position h2n, such that: h1 h1ref \u00bc I12i12\u00f0h2n h2ref\u00de; \u00f01\u00de where h1ref and h2ref are tooth reference positions, which can be set arbitrarily" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003692_978-3-642-23681-5_13-Figure13.11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003692_978-3-642-23681-5_13-Figure13.11-1.png", + "caption": "Fig. 13.11 Model of the clearance volume showing 3-axial groove", + "texts": [ + "3, we get h3 0 o2 p0 oh2 \u00fe 3 h2 0 o h0 oh o p0 oh \u00fe 1 4 D L 2 h3 0 o2 p0 o z2 K o h0 oh \u00bc 0 \u00f013:4\u00de The domain boundary conditions for the three axial grooved journal bearing are p0 \u00bc 1 at z \u00bc 0 at the grooves and elsewhere p0 \u00bc 0 and p0 \u00bc 0 at z \u00bc 1 for all h: As the hydraulic resistance is low, the lubricant will predominantly escape into the axial grooves since the inlet chamber is filled in with lubricant having a supply pressure ps. It is therefore assumed that the boundary conditions at the inlet end z \u00bc 0\u00f0 \u00de are p0 \u00bc 1 at the grooves and p0 \u00bc 0 elsewhere (Fig. 13.10). Figure 13.11 shows the model of the bearing showing three axial grooves. Reynolds boundary condition p \u00bc 0; o p=oh \u00bc 0 \u00f013:5\u00de Jakobsson-Floberg-Olsson boundary condition modified by Kicinski [33]: The JFO boundaries were applied only in land (lobe) regions, the cavitation zone is applied for the full length of the bearing as indicated in Fig. 13.12. The pressure is allowed to vary linearly in the groove from inlet to outlet. With regard to the cavitation zone it is assumed that (a) The flow in the real cavitated zone, which includes several more or less regular \u2018finger like\u2019 bubbles correspond to the flow in the theoretical zone with only one analytical \u2018bubble\u2019" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000752_detc2007-34090-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000752_detc2007-34090-Figure5-1.png", + "caption": "Figure 5\u2014Test gear installed in test fixture, cover removed for photograph.", + "texts": [ + " The film thickness at the pitch point for the operating conditions of the surface fatigue testing was calculated using the computer program EXTERN. The computing tool is an implementation of the methods of References [7] and [8]. For the purposes of the calculation, the gear surface temperature was assumed to be equal to the average of the oil inlet and outlet temperatures. Using the stated assumptions, the calculated pitch-line film thickness is 0.54 \u00b5m (21 \u00b5in). Test Procedure for Single Tooth Bending Fatigue Testing: The gear test assembly is depicted in Figure 5. The gear test specimen is positioned on a shaft which is a press fit in the fixture\u2019s casing. The test assembly was designed to conduct tests on gear teeth in sets of three. This approach was adopted in part to provide the necessary clearance for the two load rods. To permit access to the gear tooth to be tested several teeth nearby needed to be removed. Teeth were removed using the Electrode Discharge Machining (EDM) process. The upper load rod contacts the reaction gear tooth near the root of the tooth" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure4-1.png", + "caption": "Fig. 4 Cooling supply with and without setting angle", + "texts": [ + " Therefore, the geometry of the outlets had to be adapted. It was necessary to create a transition from the channel geometry inside the grinding wheel (inner geometry) to the geometry at the circumferential of the grinding wheel (outer geometry) (see Fig. 3). The coolant outlets are rotated by a setting angle relating to the centerline of the wheel. High setting angles result in a better supply of the contact zone: they increase the areas that are covered by outlets and thus by direct coolant supply (see Fig. 4). However, high setting angles are difficult to machine because of the undercuts in the transition area of the outer and inner channel geometry. As a compromise between good coolant supply and feasible undercuts, a setting angle of 45 degree has been specified. To guarantee the machinability of the undercuts, a minimal diameter of the milling tool of 3 mm is required. The resulting width of the coolant outlet is approximately 4 mm. Finally, the cross-sectional areas of the outlets and in the transition areas were held constant to assure constant flow velocities" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000698_j.clinbiomech.2007.07.017-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000698_j.clinbiomech.2007.07.017-Figure1-1.png", + "caption": "Fig. 1. Two degree-of-freedom amputated limb model. (a) Major segments and joint angles. L1\u2013L4 are length of thigh, residual shank and prosthesis; m1\u2013 m4 are mass of each segment whose centroid is located at lc1\u2013lc4, separately. (b) Schematic representation of the musculoskeletal model. Symbols appearing in the diagram are: gluteus maximus (GMAX), iliopsoas (ILIP), adductor longus (ADDL), hamstrings (HAMS), rectus femoris (RF), vasti (VAS), and gastrocnemius (GAS). (c) Schematic representation the circle cylinders representing the pathways of GMAX, ILIP, RF and VAS.", + "texts": [ + " Because the extension and flexion of the lower limb in the sagittal plane are much more important than the movements both in the frontal and transverse planes during walking, the adduction/abduction in the frontal plane and the internal/external rotation in the transverse plane of the hip and knee were ignored in this musculoskeletal modeling. The musculoskeletal model has two degree-offreedom (DoF) in the sagittal plane, in which four segments represent the prosthesis, shank, thigh, and pelvis of the amputated lower limb (see Fig. 1a). The 2-DoF was the angles of the thigh and shank relative to the horizontal line. The patella-tendon-bearing socket of the prosthesis was supposed to attach to the stump firmly without any slippage or rotation, and the solid-ankle-cushion-heel foot was simplified as a particle of mass located at the extremity of prosthesis. Segment lengths Li, segment masses mi, segment mass centers Lci, and segment inertial properties Ici (i = 1, . . ., 4) were based on the literature data (Pandy et al., 1990), especially the parameters of the shorten shank were revised by a scaling algorithm according to the subject\u2019s stump. According to the muscle positions and functions, seven independently functional muscle groups in the amputated lower limb were modeled (see Fig. 1b). The muscle groups are explained as following: (1) gluteus maximus (GMAX), (2) iliopsoas (ILIP), (3) adductor longus (ADDL), (4) hamstrings (HAMS), including semitendinosus, semimembranosus and long head of biceps femoris, (5) rectus femoris (RF), (6) vasti (VAS) including vastus lateralis, vastus intermedius and vastus medialis, (7) gastrocnemius (GAS). The three-dimensional attachments coordinate of all the muscles in the reference skeletal segments frames were based on the anthropometry data measured on the cadavers (Hoy et al", + " Especially the distal attachments of the shorten GAS in the amputating operation was changed according to the subjects\u2019 condition that is sutured via a myodesis, in which the new attachment of GAS distal end is fixed to the amputated tip of the tibial. The paths of most muscles which span the hip or the knee were approximated as straight lines, except the muscles such as HAMS, RF and so on which wrap around bones. Under these circumstances, the paths of the muscles were modeled via a cylinder separately, and the radius of the cylinders were based on the literature (Shelburne and Pandy, 1998) (see Fig. 1c). The muscle moment arms in the threedimensional space to the joints were calculated from the musculoskeletal data and the angles of the joints. The dynamic equations of motion for the model are listed according to Lagrange\u2019s laws (Pandy et al., 1990). These are expressed in the form as follows: A\u00f0q\u00de\u20acq \u00bc B\u00f0q\u00de _q2 \u00fe G\u00f0q\u00de \u00feM s\u00f0q\u00deP s \u00fe T x\u00f0q; _q\u00de; \u00f01\u00de M f\u00f0q\u00deP f \u00fe T y\u00f0q; _q\u00de \u00bc 0; \u00f02\u00de M t\u00f0q\u00deP t \u00fe T z\u00f0q; _q\u00de \u00bc 0; \u00f03\u00de where q; _q; \u20acq are vectors of thigh and shank angles, velocity, and acceleration (all are 2 \u00b7 1); A(q) is the (2 \u00b7 2) system mass matrix; B\u00f0q\u00de _q2 is a (2 \u00b7 1) vector describing both Coriolis and centrifugal effects" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001444_physreve.79.011906-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001444_physreve.79.011906-Figure5-1.png", + "caption": "FIG. 5. Forces on segment I gray rod of a fluctuating semiflexible filament. a Brownian forces: The parallel and normal Brownian forces projected in the \u2212x and \u2212y directions white rods are resolved into the corresponding resultant forces and couples. The Brownian force resultants F are written with superscripts indicating the direction of projection and segment number, and with subscripts indicating the kind of force parallel, normal, or rotational . It should be noted that the rotational force is a couple which captures the net moment of the projections of the normal Brownian forces about the center of the projection length. b Viscous forces: The viscous forces arise from the frictional resistance to motion. c Forces at cut surface: Forces V and couples d /dl latent at the cut surfaces nodes i and i+1 , arising either from the internal resistance to bending, or from the forces and couples transmitted by neighboring segments, or from the forces and couples imposed at the boundary. The lower-case superscripts indicate the node number associated with the force or couple.", + "texts": [ + " We divide the filament into segments, determine the Brownian forces on each segment, and solve a modified version of the above Euler beam equations over all segments simultaneously to balance the Brownian and other forces on each. By way of nomenclature, we denote a segment and the variables associated with it by uppercase letters, and we denote a node segment intersection and the variables associated with it by lowercase letters. A segment I is always flanked by node i on the left and node i+1 on the right. Figure 5 shows the forces and couples on a segment which has been cut off from a semiflexible filament in Brownian fluctuation. There are three kinds of forces and couples at play. 1 Brownian forces and couples: The normal Brownian forces and their role in the bending of a segment have been discussed above. The parallel Brownian forces which affect segment translation and may contribute to moments elsewhere in a curved filament need to be included in the force balance also. 2 Forces and couples at the segment interface: When a segment it cut from a filament, a remnant force V and a couple arise at the cut interface from any internal bending or stretch mechanics there, from any force or couple exerted by the neighboring segments, and from any force or couple applied externally at the filament ends" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003589_1077546311405701-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003589_1077546311405701-Figure2-1.png", + "caption": "Figure 2. Ball bearing.", + "texts": [ + "t\u00f0 \u00de M1g\u00fe FaY1 M2 \u20acX2 \u00fe C2 _X2 \u00fe K X2 X1\u00f0 \u00de \u00bc FbX2 M2 \u20acY2 \u00fe C2 \u20acY2 \u00fe K Y2 Y1\u00f0 \u00de \u00bc FbY2 M2g 9>>>= >>>; : \u00f01\u00de The ball bearing model considered here has equispaced balls rolling on the surfaces of the inner and outer races. It is assumed that the outer race is fixed rigidly to the support and the inner race is fixed rigidly to the shaft and there is no bending of races. There is perfect rolling of balls on the races so that the two points of the ball (A and B) touching the outer and inner races have different linear velocities, as shown in Figure 2, and the tangent velocity of the contact point between the ball and the inner race VA, outer race VB, respectively, can be given by VA \u00bc 0, \u00f02\u00de VB \u00bc !B RB, \u00f03\u00de at GEORGE MASON UNIV on March 2, 2013jvc.sagepub.comDownloaded from And the tangent velocity of the cage is Vcage \u00bc 1 2 VA \u00fe VB\u00f0 \u00de \u00bc VB 2 \u00bc !B RB 2 : \u00f04\u00de Because the inner race is fixed to the shaft, !B \u00bc !rotor. Then, the angular velocity of the cage is given by !cage \u00bc Vcage RA \u00fe RB\u00f0 \u00de=2 \u00bc !B RB\u00f0 \u00de=2 RA \u00fe RB\u00f0 \u00de=2 \u00bc !B RB RA \u00fe RB \u00bc !rotor RB RA \u00fe RB : \u00f05\u00de The varying compliance frequency or the ball passage frequency can be described as the cage speed times the number of balls" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001927_s12541-012-0021-7-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001927_s12541-012-0021-7-Figure11-1.png", + "caption": "Fig. 11 Worm gear efficiency measurement apparatus", + "texts": [ + " The position of the contact point and the curvature at the contact point were calculated by gear geometry and the contact area was calculated by Hertz\u2019s Law. Fig. 10 shows the calculated efficiency by the tribometer and theoretical calculations. The efficiency increases with the output torque until it reaches the peak of 90.6 % at 35 Nm. Thereafter, the efficiency starts to decrease slightly. In order to verify the efficiency prediction method and compare the theoretical efficiency with the actual efficiency of the worm gear, the efficiency measurement apparatus shown in Fig. 11 was set. A servomotor was used to apply input torque to the worm shaft and a powder brake was used to provide resistant torque to the worm wheel shaft. Two torque sensors linked to a computer measure the input and output torques. Speed of the servo motor and resistant torque of the powder brake was set by an operator via the computer. Experiments were performed in certain ranges of the output torque and speed: output torque from 2 Nm to 60 Nm and angular speed from 30 deg/s to 360 deg/s. The output torque and input torque data were acquired by the apparatus, and the efficiency of the worm gear was calculated by the following equation: 12 100 (%)out in T T m \u03b7 = \u00d7 \u00d7 (19) where, \u03b7 = efficiency of worm gear, in T = input torque, out T =output torque, 12 m = gear ratio" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001927_s12541-012-0021-7-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001927_s12541-012-0021-7-Figure2-1.png", + "caption": "Fig. 2 Force reaction on tooth surface", + "texts": [ + " The friction coefficient which is obtained from the tribometer can be mathematically converted to the worm gear efficiency by equation,2 and then the worm gear efficiency according to the normal contact pressure is obtained. In order to predict the worm gear efficiency according to the output torque, the normal contact pressure must be converted to the output torque. Gear geometry and Hertz\u2019s Law were employed for the variable conversion. When torque is applied on a worm wheel, forces act on the tooth surfaces of a worm gear. Fig. 2 shows force reactions of a worm gear mounted on 90o axes and each straddle-mounted. The force which acts on the tooth surface can be derived by simple calculations: 2 2 1 tan tan p T F R \u03b3 \u03b1= + + (3) where, F = normal load on tooth flank, T = output torque, Rp = pitch diameter of worm wheel, \u03b3 = lead angle, \u03b1 = normal pressure angle. If Rp, \u03b3 and \u03b1 are known design variables, then the relationship between output torque and normal load can be obtained from the equation. In the case of a worm gear, especially a worm gear with a plastic worm wheel, surface contact actually occurs and the pressure angle and the lead angle vary appreciably along the contact line of the surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002891_iros.2011.6094953-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002891_iros.2011.6094953-Figure2-1.png", + "caption": "Fig. 2. Coordinate frames for position-based visual servo using 4D ultrasound", + "texts": [ + " 4D Ultrasound-Based Visual Servoing We used a position-based visual servo control (PBVS) scheme described in [12] since the relative probe pose with respect to the soft tissue structures to track is calculated from the 3D rigid motion (R and t) of a target region extracted in a sequence of 3D US images. The sequential 3D US images are acquired from a 4D US probe that is mounted on the endeffector of a 6-DOF robot. We calculate the probe control velocity vc = (vc,wc) of the 4D US probe according to the 3D rigid motion of the target region where vc and wc are the translational velocity vector and the angular velocity vector of the the probe frame. We set a probe frame Fc and an object frame Fo in the center of the initial target region and the current target region, respectively shown in Fig. 2. The objective of the PBVS is to move the probe to the center of the current target region in such a way to align the probe frame Fc on the object frame Fo. The image feature s is defined as (c\u2217 tc,\u03b8u), in which c\u2217 tc and \u03b8u are a translation vector and the angle/axis parameterizations for the rotation matrix c\u2217Rc which give the coordinate of the origin and the orientation of the probe frame Fc expressed in the desired probe frame Fc\u2217 to achieve. c\u2217 tc and c\u2217Rc are given from the extracted rigid motion t and R of a target region as: c\u2217 tc =\u2212t, c\u2217Rc = R\u22121" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002213_20120914-2-us-4030.00018-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002213_20120914-2-us-4030.00018-Figure3-1.png", + "caption": "Fig. 3. The formation of agents in the presence of adversary is shown in (a). The agents shown in dashed correspond to !2. The topologies of !1 and! 2 are shown in (b) and (c), respectively. It is assumed here that the topology of! eng is similar to! 1 in the sense that Leng = L, where L is the Laplacian associated with! 1.", + "texts": [ + " The last statement is an immediate corollary of Lemma 5. ! Our assumption on P2 is conservative and the result likely holds for *P2 Hurwitz, since *V %1 Leng P2VLeng is also Hurwitz and thus can potentially satisfy the conditions of Theorem 1. Also, note that agents in\" 2 can detect the presence of the adversary by choosing $ = 0 and evaluating the changes in the estimates of\" 1 received from their neighbors in this network. 6. SIMULATIONS Consider a network\" 1 with five agents {v1, . . . , v5} as shown in Figure 3. Suppose these agents wish to execute a kinematic-based flocking algorithm (see [Lee and Spong, 2007]), such that they achieve a formation in which xj * xi = j * i, for all i, j ! {1, . . . , 5} (this flocking position is consistent, in the sense of [Lee and Spong, 2007]). The agents can communicate their positions with each other according to the topology shown in Figure 3(b). Since this topology is connected, a consensus dynamics can be used to achieve formation. Now suppose that there is an adversary that can influence the estimates of each agent about the states of its neighbors, according to (4). We consider two cases: (i) P2 = 3I5 and P1 = * 2 2 2 + 1.7730 1.4254 0.0849 1.6351 1.9459 0.0573 1.0009 0.1429 1.4449 1.2980 0.9798 0.9422 1.0433 0.2997 1.6007 0.3359 0.1192 0.1935 1.3192 0.9076 1.9574 1.3639 1.6363 1.0372 0.8648 , 3 3 3 - ; (ii) P2 = I5 and P1 = * 2 2 2 + 3", + " Figures 4 (a) and (b) show the impact of the signals of the adversary on the stability of the formation dynamics for each case, respectively. In the first scenario, the dynamics of the interconnected system (\"1,\"2) is asymptotically stable, however, its equilibrium does not correspond to the equidistance formation position without the presence of the adversary. In the second scenario, however, the presence of the adversary causes instability. Suppose a second (virtual) network\" 2, with topology shown in Figure 3(c), is interconnected with\" 1, where the topology of\" eng is such that Leng = L. The interconnected system (\"1,\"adv,\"2) is then asymptotically stable, as shown in Figure 4. In (a), for $ = 10, the final relative position of the agents are given by x2 * x1 = 1.0085, x3 * x2 = 0.9926, x4 * x3 = 1.0095, x5 * x4 = 0.9917, which is significant improvement over the one in the presence of adversary and without\" 2: x2 * x1 = 2.1572, x3 * x2 = 0.3947, x4 * x3 = 1.8782, x5 * x4 = *0.4383. In (b), for $ = 10, the final relative positions of the agents are given by x2 * x1 = 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000559_j.triboint.2007.07.009-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000559_j.triboint.2007.07.009-Figure1-1.png", + "caption": "Fig. 1. ASTM G98 button-on-block test configuration.", + "texts": [ + " The ASTM G98 test method was presented by Schumacher of the Armco Steel Corporation [13] in the early 1970s and has been used since that time as a way to determine the relative galling resistance of one material pair versus another. The results of the test do not yield quantitative data that can be used for design purposes; however, the results can be used to rank the material pairs in terms of their galling resistance. The design engineer can then choose the material couple with the least likelihood of galling. The ASTM G98 Standard Test Method for Galling Resistance of Materials, often referred to as the button-onblock test, calls for a cylindrical button specimen to be mated to a block specimen as shown in Fig. 1. A constant load is placed on the button with a universal testing machine and it is then rotated one revolution relative to the block. Upon completion of the test, the specimens are examined visually without magnification for the presence of galling. If galling has not occurred, then new specimens are loaded at a higher level and tested. This is continued until galling occurs, at which point the load of the highest non-galled test is averaged with the load of the galled test. This load is divided by the cross-sectional area of the button specimen and is deemed the threshold galling stress" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002740_1.4823475-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002740_1.4823475-Figure1-1.png", + "caption": "Fig. 1. A ball with center-of-mass coordinate r rolls without slipping on a cylindrically symmetric surface given by z\u00f0q\u00de; only the vertical slice at angle u is shown. Note that dz=dq\u00bc z0 \u00bc tan h is the slope of the surface\u2019s tangent line in the radial direction.", + "texts": [ + " We seek to sharpen the approach taken there and obtain equations of motion that account for the full vector nature of the usual rolling-without-slipping constraint. Thus, we start with the vector form6 of Eq. (1) rendered in cylindrical coordinates v \u00bc ax n\u0302 \u00bc \u00f0 _q;q _u; _z\u00de \u00bc \u00f0axu cos h; axq cos h axz sin h;axu sin h\u00de; (2) where n\u0302 is the normal to the surface of revolution (with symmetry axis aligned with the z-direction) and h is the angle between horizontal and the radial tangent to the surface at the point of contact (see Fig. 1). We will find it useful to reexpress Eq. (2) in terms of the angular and radial components of the spin as xu \u00bc _q a cos h (3) and xq \u00bc L qa cos h \u00fe xz tan h ; (4) where L \u00bc q2 _u (5) 890 Am. J. Phys. 81 (12), December 2013 http://aapt.org/ajp VC 2013 American Association of Physics Teachers 890 This article is copyrighted as indicated in the article. Reuse of AAPT content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to IP: 132.177.211.166 On: Fri, 04 Apr 2014 20:13:14 is the orbital angular momentum per unit mass" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002293_macp.201200239-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002293_macp.201200239-Figure2-1.png", + "caption": "Figure 2 . Experimental setup for IR dichroism measurements and polarizing optical microscopy.", + "texts": [ + " After the reaction, the glass substrates were removed from the cell, and the resultant gel fi lms were left to swell in dichloromethane in order to wash away the unreacted materials and the nonreactive solvent. The swollen gels were gradually deswollen by adding methanol to the swelling solvent, and the fully dried fi lms were employed as the specimens. The temperatures T NI of the I-PNE and N-PNE were evaluated to be 110 and 111 \u00b0 C by polarizing optical microscopy, respectively. The IR dichroism measurements and texture observations of the specimens under stretching were performed at 60 \u00b0 C in the nematic state using a custom-built stretching device (Figure 2 ). www.MaterialsViews.com l page numbers, use DOI for citation !! DOI: 10.1002/macp.201200239 bH & Co. KGaA, Weinheim Memory and Development of Textures . . . www.mcp-journal.de Macromolecular Chemistry and Physics Rectangular fi lm specimens were clamped at both ends, and the effective dimensions of the samples excluding the clamped area were 5 mm \u00d7 1 mm \u00d7 32 \u03bc m. The clamps could be shifted by micrometer-sized screws, so that the imposed strain could be controlled with high precision. The device could stretch the specimens equally from both ends so that the IR beam, which was focused at the center of the specimens and whose spot was 100 \u03bc m \u00d7 100 \u03bc m in size, could be made to remain focused at the same position during the stretching" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003931_j.procir.2014.02.038-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003931_j.procir.2014.02.038-Figure3-1.png", + "caption": "Fig. 3. In yellow, support structures under an overhang surface.", + "texts": [ + " For the production of functional parts, these deformations may not be acceptable. Indeed, they can be out of the expected tolerance and will require a surface finishing by cutting process for example (when it is possible). Moreover, these deformations are the result of internal stresses; it is possible that there are residual stresses in the final part. That is why reducing the curling effect is a real issue that has to be addressed to produce complicate parts and especially horizontal planes. The current solution to solve this problem is the use of support structures (Fig. 3). By different ways, these structures allow to reduce the curling effect. These structures are removed after the build and because of their material price, the consumed time to remove the supports and the accessibility problems, this solution cans not always be use. That is why the aim of this study is to explore another solution: the modulation of the energy that is used to melt the powder. In this paper, the link between the conductivity of the material and the overheating of negative surface will be analysed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000940_978-3-540-70701-1_3-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000940_978-3-540-70701-1_3-Figure2-1.png", + "caption": "Fig. 2. Relevant camera and object frames", + "texts": [ + " For eye-to-hand cameras, the matrix Hc is constant, and can be computed through a suitable calibration procedure, while for eye-in-hand cameras this matrix depends on the camera current pose xc and can be computed as: Hc(xc) = He(xe)eHc where He is the homogeneous transformation matrix of the end effector frame e with respect to the base frame, and eHc is the homogeneous transformation matrix of camera frame with respect to end effector frame. Notice that eHc is constant and can be estimated through suitable calibration procedures, while He depends on the current end-effector pose xe and may be computed using the robot kinematic model. The relevant frames and the transformation matrices are illustrated in Fig. 2, where the more general case of multiple mobile and fixed cameras is depicted. Therefore, the homogeneous coordinate vector of P with respect to the camera frame can be expressed as cp\u0303 = cHo(xo,xc)op\u0303 (5) where cHo(xo,xc) = cH\u22121(xc)Ho(xo). Notice that xc is constant for eye-tohand cameras; moreover, the matrix cHo does not depend on xc and xo separately but only on the relative pose of the object frame with respect to the camera frame. The velocity of the camera frame with respect to the base frame can be characterized in terms of the translational velocity of the origin vOc and of angular velocity \u03c9c" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000253_978-1-4020-8600-7_33-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000253_978-1-4020-8600-7_33-Figure2-1.png", + "caption": "Fig. 2 Transmission angle of the 3-RRR manipulator.", + "texts": [ + " For instance, the previous concept was used in [5] to analyze the kinetostatic performance of manipulators with multiple inverse kinematic solutions, and therefore to select their best working mode. The transmission angle can be used to assess the quality of force transmission in mechanisms involving passive joints. Although it is well known and easily computable for 1-DOF or single loop mechanisms [3, 8], it is not extensively used for n-DOF mechanical systems (n > 1) [2]. The transmission angle \u03c8i is defined as an angle between vectors of force Fci and translational velocity Vci of a point to which the force is applied as illustrated in Figure 2. When link AiBi is driven, the direction of force Fci is the direction of link BiCi , namely, \u03b3i = arctan ( yCi \u2212 yBi xCi \u2212 xBi ) , i = 1, 2, 3 (10) 315 N. Rakotomanga et al. Conversely, when link AiEi is driven, the direction of force Fci is the direction of line (AiCi), namely, \u03b3i = arctan ( yCi \u2212 yAi xCi \u2212 xAi ) , i = 1, 2, 3 (11) The instantaneous centre of rotation depends on the leg under study. For example, instantaneous centre of rotation I1 associated with leg 1 is the intersecting point of forces Fc2 anf Fc3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000262_b712482a-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000262_b712482a-Figure1-1.png", + "caption": "Fig. 1 a: The polymerization flow cell consists of a quartz plate with access holes, a parafilm spacer, a gel support film, and a solid quartz plate. b: The degassed CCA prepolymer suspension is injected through tubing connectors and sealed from the environment in order to exclude oxygen during polymerization.", + "texts": [ + " The diffraction preak of the CCA was measured using a 6 around 1 reflectance probe and a fiber-optic diode spectrometer with a tungsten halogen light source (Ocean Optics). PCCA were formed by photopolymerization of acrylamide (0.10 g, 1.4 mmol, Sigma) and N,N \u2032-methylenebisacrylamide (0.002 g, 13 lmol, Fluka) dissolved in the CCA (20% w/w dispersion of polystyrene latex spheres) following the procedures of Asher et al.28 Before polymerization, the colloidal suspension and prepolymer solution was degassed, and the evacuation chamber was backfilled with nitrogen. This solution was injected into a polymerization cell (Fig. 1) that was evacuated in order to remove oxygen. A gel support film containing surface vinyl groups (Bio-rad) was used as one face of the polymerization cell, covalently attached to the PCCA during photopolymerization (12 min with two 100 W UV lamps, Black Ray). After polymerization the PCCA was removed from the polymerization cell and cut into rectangular pieces (approximately 8 mm \u00d7 12 mm) with a razor blade and stored in deionized water. Metal ion sensors were made from the PCCA by the method of Asher et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002591_10402004.2011.626144-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002591_10402004.2011.626144-Figure6-1.png", + "caption": "Fig. 6\u2014Effect of roughness orientation on journal center trajectory.", + "texts": [ + " As the fluid becomes elastic, the operating value of eccentricity increases and hence the minimum film thickness is reduced in the case of a viscoelastic fluid. Further, since the transformed external load component F\u0304\u2217 ex leads the load F\u0304ex as shown in Fig. 2, it attains the value of a load that is required to produce a maximum value of minimum film thickness at a smaller crank angle and, consequently, the position of the maximum value of h\u0304min shifts toward the left (i.e., toward smaller crank angle). Figures 6 to 11 show the journal center trajectories during the second load cycle. Figure 6 shows a comparison of the journal center locus of a smooth bearing with the journal center locus of a rough bearing that has two-sided roughness. The rough bearing traces almost the same orbit as that of the smooth bearing but the rough bearing, especially with a longitudinal roughness pattern, maintains a more enhanced film thickness than that of the smooth bearing during the entire load cycle. As seen from Fig. 7, the bearing with stationary roughness maintains larger film thickness throughout the load cycle compared to the two-sided and moving roughness cases but its journal center traces a larger orbit" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001895_978-3-642-32448-2_11-Figure11.5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001895_978-3-642-32448-2_11-Figure11.5-1.png", + "caption": "Fig. 11.5 Discretized chip geometry", + "texts": [ + "tooth tooth int dt dt n = In every time step ti=1,\u2026, nint the material removal is calculated, where 2tooth toothdt is the tooth-passing time calculated by 2 60 \u00b7tooth tooth z dt N n = with the number of revolutions n and the number of teeth .zN The material removal simulation calculates a point cloud representing the outer volume of a single chip, which the discrete chip geometry is extracted from. The chip geometry is described by the angular discretized chip thickness h(\u03c6,z) with the entry and exit angles \u03c6in, \u03c6out. As the chip thickness varies along the cutting edge the chip gets subdivided into discs of the height dz and in d\u03c6 in angular direction (cf. Fig. 11.5). The chip thickness is calculated in three steps. In the first step, the set of points is subdivided into disks of the height dz. As a next step, inside each disc the maximum angular points are defined as restricting points by \u03c6in and \u03c6out (cf. Fig. 11.6a). Due to the complex contact conditions and the complex chip geometries the entry and exit angles vary over the depth of cut ap. The cross-sectional geometry of each disk is calculated based on the cylindrical form of the cutter at the discrete time of ti and ti-1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001795_s11012-010-9319-7-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001795_s11012-010-9319-7-Figure1-1.png", + "caption": "Fig. 1 Geometry of a worn and misaligned journal bearing", + "texts": [ + "gr ment provides a tool for optimally reducing the parameter space of independent variables in EHD journal bearing design. Keywords Journal bearings \u00b7 EHD lubrication \u00b7 Wear \u00b7 Misalignment \u00b7 Pareto optimization Nomenclature A Bearing interior surface Aj Journal surface c Radial clearance e0 Bearing eccentricity E Young\u2019s modulus of elasticity f = Ffr/Fhydr Friction coefficient f\u0304 = f Rj/c Normalized friction coefficient Ff r = \u222b\u222b Aj \u03c4dAj Friction force on the journal surface Fhydrx = 0 x component of hydrodynamic force at equilibrium (Fig. 1) Fhydry = \u222b\u222b Aj pdAj = \u2212W y component of hydrodynamic force at equilibrium (Fig. 1) h Film thickness L Bearing length N Rotational frequency of journal Ob,Oj Bearing, journal centre p Hydrodynamic pressure p\u0304 = (p \u2212 pa)c 2/\u03bc\u03c9R2 j Dimensionless pressure pa Ambient (reference) pressure Rj ,D = 2Rj Journal radius, diameter S = \u03bcNDL(Rj/c) 2/Fhydr Sommerfeld number th Bearing pad thickness (Fig. 1) U Journal surface velocity vector W Bearing external load Greek symbols \u03b40, d0 = \u03b40/c Maximum, and corresponding dimensionless wear depth (Fig. 1) \u03bc Lubricant dynamic viscosity \u03c60 Attitude angle \u03c80 = \u03bc\u03c9/E(th/Rj )(Rj/c) 3 Bearing deformation coefficient \u03c8x,\u03c8y Misalignment angles: yz plane, xz plane (Fig. 1) \u03c8x = \u03c8xL/c,\u03c8y = \u03c8yL/c Normalized misalignment angles: yz plane, xz plane \u03c9 Angular velocity of journal Misalignment and wear are two of the most important faults of journal bearings, with misalignment being also a primary cause of wear. Misalignment can also be a particular working bearing condition, due to the deflections of the shafting system. Proper shafting support and alignment reduce the effect of bearingshaft misalignment; a misalignment angle less than 0.3 mrad is generally acceptable in most industrial and marine applications [1]", + " Here, we solve numerically the elastohydrodynamic problem, and formulate a minimization problem, with the normalized friction coefficient (associated with bearing performance) and the normalized maximum pressure (associated with bearing strength) as objective functions. Optimal solutions are identified using the concept of Pareto dominance, and are physically interpreted. An important outcome of the present work is an engineering tool for determining optimal design parameters of journal bearings of different health states, operating in the elastohydrodynamic regime. In the present work, the bearing is considered to be elastic, and the flow steady and isothermal. The geometry of the worn bearing follows the model of [5], and is shown in Fig. 1; here, Ob is the bearing centre, Oj the journal centre, Rb the bearing radius, Rj the journal radius, e0 the bearing eccentricity at equilibrium, \u03c60 the attitude angle, and L the bearing length. The external load W is assumed vertical (i.e., along the y-axis) and constant. The Reynolds equation, (1), is solved for the pressure distribution using finite element techniques. \u2207 \u00b7 ( h3 6\u03bc \u2207p ) = \u2207 \u00b7 (h U) (1) where U is the velocity vector of the journal surface, with | U | = \u03c9Rj ,p(\u03b8, z) the pressure distribution, h(\u03b8, z) the film thickness, and \u03bc the dynamic viscosity of the fluid", + " The Reynolds boundary condition, not utilized here, assumes that the positive pressure curve terminates with a zero gradient in the divergent part of the film; it gives in some cases more accurate results than the half Sommerfeld boundary condition. Nonetheless, it is still an approximation to the transition from singlephase flow to multi-phase flow, and is computationally more costly. Following the abrasive wear model of [5], the worn zone of the bearing pad forms an arc, which is centered around the vertical (load) direction, is uniform in width along the bearing length, and does not include smaller scale wear areas (see Fig. 1). This model is based on the hypothesis that the bearing pad wears out due to the footprint created by the shaft to the pad; the worn arc has a radius larger than that of the journal. The change in the film geometry due to wear of the bearing pad is given by (3): \u03b4h(\u03b8) = \u23a7 \u23aa\u23a8 \u23aa\u23a9 c(d0 \u2212 1 \u2212 cos(\u03b8 + \u03c60)) for \u03b8w1 \u2264 \u03b8 \u2264 \u03b8w2 0 for \u03b8 < \u03b8w1, or \u03b8 > \u03b8w2 (3) where d0 is the dimensionless maximum wear depth (ratio of the wear depth, \u03b40, to the bearing radial clearance, c). Angles \u03b8w1 and \u03b8w2 define the extent of the footprint on the bearing pad (see Fig. 1), and can be determined from (4), for given dimensionless wear depth d0: cos(\u03b8 + \u03c60) = d0 \u2212 1 (4) The hydrodynamic film thickness for the worn and misaligned bearing is calculated from the film thickness corresponding to the intact bearing, suitably modified to take into account the journal misalignment (in the yz and xz planes), and the geometry of the worn region, see Fig. 1. Quantitatively, (5) is obtained for the film thickness [2]: hhydro(\u03b8, z) = c + e0 cos(\u03b8) \ufe38 \ufe37\ufe37 \ufe38 Intact bearing + z[\u03c8y cos(\u03b8 + \u03c60) + \u03c8x sin(\u03b8 + \u03c60)] \ufe38 \ufe37\ufe37 \ufe38 Modification for misalignment + \u03b4h(\u03b8) \ufe38 \ufe37\ufe37 \ufe38 Modification for wear (5) From dimensional analysis of the EHD problem, it follows that the physics depends on the length-todiameter ratio, L/D, the Sommerfeld number, S, and the bearing deformation coefficient, \u03c80. In the present work, the bearing pad is considered elastic, and is housed inside a rigid bearing shell (see Fig. 1). The deformable bearing pad is modeled with two layers of linear 3-D wedge elements featuring 6 nodes per element, see Fig. 2. The bearing pad is assumed to be simply supported along the exterior surface, i.e., the local displacement vector is taken {d} = 0. The element faces in contact with the lubricant film are subjected to normal loads due to the hydrodynamic pressure, calculated by (1). The corresponding bearing pad elasticity problem is solved with the ABAQUS FEM software [16]. We note that the present model can be also applied to the problem of a rigid bearing pad featuring a thick deformable coating in the interior pad surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003076_iros.2012.6386127-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003076_iros.2012.6386127-Figure2-1.png", + "caption": "Fig. 2. Representation of precision and set tasks", + "texts": [ + " As a matter of fact the introduction of dynamic programming originally appeared within different works of part of the authors; for instance, the distributed control of modular robotic structures, seen as an aggregation of 1- d.o.f. micro-units, is dealt with in [4]. This approach was then extended to the more general case of robots composed by macro-units, as the arm and the vehicle, within the first seminal work [3] related with the TRIDENT project. In the TRIDENT terminology, all the inequality and equality control objectives to be achieved are denoted as set and precision tasks, respectively. The detected set objectives for the floating manipulation system (roughly sketched in Fig. 2) are: (a) joint limits: each arm joint within its range; (b) manipulability: arm scalar manipulability measure above a given positive threshold; (c) horizontal attitude: module of misalignment angle be- tween vehicle z- axis and the absolute vertical line, below a given threshold; (d) camera centering: module of the misalignment angle between the (fixed to the vehicle) camera axis and the target line of sight, below a given threshold; (e) camera height: scalar height of the camera with respect to the target on the sea floor, above a given threshold; (f) camera distance: scalar distance of the camera from the absolute vertical line passing through the target, below e given threshold. On the other hand, only the end-effector approach (modules of the end-effector vector distance and attitude with respect to the target to be both zeroed) has been identified as precision objective. Figure 2 exemplifies two of these tasks, namely the end-effector approach and the camera centering, together with their representative error vectors. To solve the control problem, the system must be driven in a total concurrent way toward a feasibility set, in which all the required objectives are achieved without conflicts. To this aim, the employment of closed-loop priority-based concurrent tasks plays the leading role, concurrently integrating higher priority objectives with lower priority ones to obtain the desired robot behavior in a faster and agile way" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003447_iecon.2012.6389263-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003447_iecon.2012.6389263-Figure3-1.png", + "caption": "Fig. 3. Artificially deteriorated bearings: (a) outer race deterioration, (b) inner race deterioration, (c) cage deterioration, (d) ball deterioration.", + "texts": [ + " Indeed, it has been mentioned in a number of previously published paper, that one of the main difficulties in real word testing of developed condition monitoring technique, is the lack of collaboration needed with wind turbine operators and manufacturers, due to data confidentiality, particularly when failures are present [3]. In this paper, the authors propose a low complexity detector which does not require any training sequence. Indeed, the proposed detector is based on the dominant imfvariance. III. EXPERIMENTAL EVALUATION OF THE EEMD-BASED FAILURE DETECTION ApPROACH The failures are obtained by simply drilling holes in different parts of the bearing (Fig. 3). A conventional test bed is used in order to test the proposed failure detection method (Fig. 4). The mechanical part (Fig. 7a) is composed by a synchronous and an induction machine. The induction machine is fed by the synchronous generator in order to eliminate time harmonics. Indeed, this will automatically eliminate supply harmonics and therefore allow focusing only on bearing faults effect on the stator current. The tested induction machine has the following rated parameters: 0.75 kW, 220/380 V, 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003324_10402004.2011.604758-Figure16-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003324_10402004.2011.604758-Figure16-1.png", + "caption": "Fig. 16\u2014Numerical reconstruction of the experimental film temperature field. Surface map for one land of the bronze test bearing land for the 0.1 MPa LS (rel) tests. (color figure available online.)", + "texts": [], + "surrounding_texts": [ + "The operating conditions were selected to observe the influence of supply pressure on a lightly loaded CGJB and to determine whether the method of measurement could be implemented on the Institut Pprime test rig with good results. In addition, the extensive measurements provided a good basis for checking numerical models, because there is a shortage of detailed studies in the literature. The current experiment represents a first step toward a broad chain of tests aimed to monitor the bronze CGJB in various operating conditions for typical clearances. Prior to presenting and commenting on the results, the following remarks are necessary. The method of rotating the CGJB to collect extensive data with a limited number of sensors is not novel. Furthermore, it is impossible to obtain identical measurements for any test even though the modus operandi, sensors, or operating conditions are similar. There is always a certain amount of uncertainty in all experiments, and mechanical\u2013electronic conversions are only a small contribution to the overall uncertainties. Furthermore, heat flows through conduction: from the housing Fig. 6\u2014Circumferential pressure distributions for the 0.1 MPa LS (rel) tests at one land of the bronze bearing. D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 Fig. 7\u2014Circumferential pressure distributions for the 0.2 MPa LS (rel) tests at one land of the bronze bearing. through the hydrostatic bearings down to the pneumatic cylinder; from the shaft throughout the entire test rig; and from the shaft and housing to the environment through natural convection. These effects characterize all experimental measurements, but the effect is high before thermal equilibrium is reached. However, it is important to limit these uncertainties and to be aware of them. The authors consider that combining the test results is adequate, because the tests were performed in very good conditions. For this reason, test results are presented graphically and provide a good basis for drawing our conclusions. Tables are given as support to highlight some of the uncertainties and to provide supplemental information on how the tests were performed and how the data have been processed. Figures 6\u201311 contain information on pressure, Figs. 12\u201317 on temperature, and Figs. 18 and 19 present correlations between pressure and temperature measurements. Figures 6 and 7 and 12 and 13 present combined measurements obtained from all 13 bearing rotation cases for both supply pressures. The coordinates system is the same as in the previous sections. At first glance, the supply pressure influences both the values and locations of pressure and temperature peaks. For a higher supply pressure, the pressure and temperature peaks are shifted downstream. Also, a higher supply pressure raised the minimum film pressure from a value close to 0.02 MPa to about 0.035 MPa. The measured mid-groove pressure is not been superposed in Figs. 6 and 7 because its value remained constant. That was not D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 the case for the mid-groove temperature values, because their values being much higher than the 40\u25e6C supply temperature. Figures 8 and 9 and 14 and 15 represent numerical reconstructions of the measured values into 2D contour maps. For all reconstructions, pressure or temperature, the grid was created using the natural neighbor method. The contours and their corresponding labels correspond to constant pressure isobars (MPa) for Figs. 8 and 9. Similarly, constant temperature isotherms are drawn (\u25e6C) for Figs. 14 and 15. The same line of reasoning is maintained for Figs. 18 and 19. The pressure grid contains measurements performed at the five axial planes and boundaries. The environment boundary consists of three additional planes, at \u22122, \u22125, and \u221230 mm, with average atmospheric pressure measured during the tests. Correspondingly, the supply groove boundary consists of three additional planes, at 21, 24, and 30 mm, with average pressure measured in the mid-groove during the tests. The axial grid spacing is 3 mm and the circumferential step size is 10\u25e6. The temperature grid contains measurements performed at the four axial planes and boundaries. The environment boundary consists of two additional planes, at \u22125 and \u221230 mm, with the average environment temperature measured during the tests. Correspondingly, the supply groove boundary consists of two additional planes, at 25 and 30 mm, with the temperatures measured in the mid-groove during the tests. The axial grid spacing is 6 mm and the circumferential step size is 10\u25e6. Choosing two axial grid steps for the plotting results was necessary due to the different arrangement for the measurement planes on the land and physical considerations regarding the boundaries. One may expect the pressure outside the bearing to be at atmospheric value all around the circumference. This is not the case for the temperatures, and because there was no possibility to measure temperature all around the circumference, the boundary was limited to two planes. Grid construction with just one plane, far outside, was not sufficient. The grids constructed for each of the cases presented in Figs. 8 and 9 and 14 and 15 are maintained for the 3D numerically reconstructed surface maps in Figs. 10 and 11 and 16 and 17 and for the 2D contour map comparisons in Figs. 18 and 19. The surface maps corresponding to the pressure fields (Figs. 8 and 9) are plotted within the 0 to 26.5 mm axial interval. This enables observing the smooth pressure distribution in the grooves and the slightly higher pressure values at the environment outlet. The temperature field surface maps (Figs. 14 and 15) are displayed in two views corresponding to the same land. The alternate view represents the main view seen from the opposing corner. The same axial limitation is imposed on plotting the results as the case for pressure. Extending the plot toward the environment would have shown only abrupt temperature drops. Considering the pressure plots, downstream of the sharp pressure drop region, the film pressure reaches near atmospheric values after about 30\u201350\u25e6, depending on supply pressure. Because D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 Fig. 10\u2014Numerical reconstruction of the experimental film pressure field. Surface map for one land of the bronze test bearing land for the 0.1 MPa LS (rel) tests. (color figure available online.) the temperature values did not increase in the low-pressure region, downstream of the sharp pressure drop, it is safe to assume that air is entrained from the environment and that lubricant viscous shear is minimal. Taking into account that the environment supply temperature boundary can be a bit rough, one may be surprised to observe that the temperature distributions presented in Figs. 14 and 15 have an axisymmetrical characteristic. This is not the case for the pressure distributions, where the Poiseuille component shifts the isobars axially; that is, pressure distributions are asymmetrical with respect to the mid-land plane. This can be also seen at the 3D surface plots in Figs. 10 and 11 and 16 and 17 or directly on the superposed pressure and temperature contour plots in Figs. 18 and 19. Superposing the pressure and temperature fields for the lightly loaded CGJB in steady-state conditions, Figs. 18 and 19 reveal that maximum temperatures are encountered in the same region of the sharp pressure zone. As Coyne and Elrod (36), (37) explained, the film detaches radially from the stationary boundary, that is, the bushing, and continues to be spread by the shaft Fig. 11\u2014Numerical reconstruction of the experimental film pressure field. Surface map for one land of the bronze test bearing land for the 0.2 MPa LS (rel) tests. (color figure available online.) D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 Fig. 12\u2014Circumferential temperature distributions for the 0.1 MPa LS (rel) tests at one land of the bronze bearing. Fig. 13\u2014Circumferential temperature distributions for the 0.2 MPa LS (rel) tests at one land of the bronze bearing. Fig. 14\u2014Numerical reconstruction of the experimental film temperature field. Contour map for the 0.1 MPa LS (rel) tests at one land of the bronze bearing. D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 but with a lower film thickness. Hence, if the low-thickness film moves with an almost constant shaft speed, viscous shear would be minimal or even zero. The remaining gap is filled with air, because the CGJB is open to the environment, and vaporous cavitation is characteristic of dynamic operating regimes. It is worthwhile mentioning that for a fully immerged journal bearing (Braun and Hendricks (9)), the cavitation zone is filled with airlike gases. Note that the film recovers in both axial and circumferential directions; that is, it is not sudden. Measured atmospheric pressure values were slightly higher than 0.1 MPa. Table 3 summarizes the measurements performed at z = 7 mm the 0.2 MPa LS (rel) tests. The first measurements were performed with the bearing positioned at 15 o\u2019clock; that is, a \u221260\u25e6 shift, called m60, shown in Table 3. The bearing was rotated sequentially with a 10\u25e6 step, down to 9:30 o\u2019clock, or a +50\u25e6 shift, called p50. For this z = 7 mm plane and current operating conditions, it was not necessary to perform many pressure sensor changes. However, this is not the case if operating conditions are changed. The standard deviations corresponding to the pressure values tended to increase with an increase of the average value. This is valid, because uncertainties for these sensors increase with the measured values. Another possible reason is a slight influence of the supply temperature fluctuations on the bearing geometry, that is, small expansions and contractions, or changes in lubricant viscosity. Recall that the supply temperature variation is slow. A period lasts approximately 10 min for the 0.2 MPa LS (rel) tests and approximately 15 min for the 0.1 MPa LS (rel) tests. The film temperatures also followed a periodical variation, very similar to that of the supply temperature, but with a phase delay possibly due to thermal inertia. Still, all measurements were taken over long time intervals, enabling fluctuations to be averaged to the values corresponding to 40\u25e6C supply temperature. The load was maintained very precisely for both test cases, with only a couple of N standard deviation, as seen in Tables 4 and 5. Note that even if the proximity probes detected some variations in the shaft locus, they were about 2 \u00b5m for the 0.1 MPa LS (rel) tests and about 1 \u00b5m for the 0.2 MPa LS (rel) tests. This contributes to the conclusion that the higher supply pressure test was more stable, because all of the temperature values measured revealed a smoother distribution than the lower supply pressure tests. Operating eccentricity and attitude angle values for the lowload test cases were determined relative to high-load reference tests, summarized in Table 6, by considering shaft locus variation with respect to the reference tests (Table 7) and adequate numerical models. With respect to the high-load reference tests, the D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 Fig. 18\u2014Numerical reconstruction of the experimental film pressure and temperature fields. Contour map comparison for the 0.1 MPa LS (rel) tests at one land of the bronze bearing. Fig. 19\u2014Numerical reconstruction of the experimental film pressure and temperature fields. Contour map comparison for the 0.2 MPa LS (rel) tests at one land of the bronze bearing. D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 TABLE 3\u2014TEST CASE 0.2 MPA LS (REL)\u2014PRESSURE AND TEMPERATURE MEASUREMENTS AT Z = 7 MM Pressure Temperature Measurements Measurements Theta (\u25e6) Average (MPa) (abs) SD (MPa) (abs) Sensor Type (S.T.) Shift Bearing Position (o\u2019clock) Average (\u25e6C) SD (\u25e6C) S.T. Shift Bearing Position (o\u2019clock) 0 0.1533 0.0021 A m10 12:30 43.1 0.9 K m40 14 10 0.1609 0.0012 A 0 sh 12 43.4 0.8 K m30 13:30 20 0.1654 0.0006 B p10 11:30 43.1 0.7 K m20 13 30 0.1674 0.0004 B p20 11 43.1 0.7 K m10 12:30 40 0.1703 0.0003 B p30 10:30 43.2 0.8 K 0 sh 12 50 0.1712 0.0002 B p40 10 43.3 0.7 K p10 11:30 60 0.1734 0.0002 B p50 9:30 43.1 0.6 K p20 11 70 0.1760 0.0002 B m60 15 43.2 0.7 K p30 10:30 80 0.1813 0.0002 B m50 14:30 43.2 0.6 K p40 10 90 0.1852 0.0005 B m40 14 43.3 0.8 K p50 9:30 100 0.1962 0.0007 B m30 13:30 43.8 0.8 K m60 15 110 0.1998 0.0006 B m20 13 43.9 0.8 K m50 14:30 120 0.2046 0.0008 B m10 12:30 43.8 0.9 K m40 14 130 0.2158 0.0011 B 0 sh 12 44.5 0.8 K m30 13:30 140 0.2466 0.0010 C p10 11:30 44.6 0.7 K m20 13 150 0.2924 0.0011 C p20 11 44.7 0.8 K m10 12:30 160 0.3259 0.0012 C p30 10:30 45.2 0.9 K 0 sh 12 170 0.4234 0.0018 C p40 10 45.7 0.7 K p10 11:30 180 0.5658 0.0056 C p50 9:30 45.9 0.6 K p20 11 190 0.7990 0.0064 C m60 15 46.2 0.8 K p30 10:30 200 1.0925 0.0166 C m50 14:30 46.6 0.6 K p40 10 210 1.2310 0.0092 C m40 14 46.8 0.8 K p50 9:30 220 0.9179 0.0112 C m30 13:30 46.7 0.7 K m60 15 230 0.2049 0.0272 C m20 13 46.7 0.6 K m50 14:30 240 0.0454 0.0004 C m10 12:30 46.2 0.8 K m40 14 250 0.0654 0.0008 C 0 sh 12 46.2 0.6 K m30 13:30 260 0.0781 0.0011 A p10 11:30 45.8 0.5 K m20 13 270 0.0930 0.0002 A p20 11 45.2 0.5 K m10 12:30 280 0.0965 0.0007 A p30 10:30 45.1 0.6 K 0 sh 12 290 0.0994 0.0003 A p40 10 45.0 0.5 K p10 11:30 300 0.1030 0.0004 A p50 9:30 44.7 0.4 K p20 11 310 0.1017 0.0002 A m60 15 44.5 0.5 K p30 10:30 320 0.1030 0.0002 A m50 14:30 44.5 0.4 K p40 10 330 0.1039 0.0001 A m40 14 44.1 0.6 K p50 9:30 340 0.1251 0.0036 A m30 13:30 43.9 0.6 K m60 15 350 0.1359 0.0028 A m20 13 43.6 0.7 K m50 14:30 D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 TABLE 5\u2014TEST CASE 0.2 MPA LS (REL) \u2013 OPERATING CONDITIONS Load Proximity Probe 1 Proximity Probe 2 Supply Temperature Speed Env. Temp. Shift Ave. (N) SD (N) Ave. (rpm) Ave. (V) SD (V) Ave. (V) SD (V) Ave. (\u25e6C) SD (\u25e6C) Ave. (\u25e6C) m60 997 6 1,001 4.0771 0.0092 4.7554 0.0091 39.9 0.7 34.2 m50 996 7 1,001 3.9924 0.0091 4.9073 0.0093 40.1 0.6 35.0 m40 1,005 10 1,000 4.0515 0.0085 4.8725 0.0101 39.9 0.8 35.2 m30 1,000 4 1,001 4.1540 0.0106 4.7570 0.0100 39.9 0.7 35.7 m20 1,000 4 1,001 4.0848 0.0080 4.8392 0.0098 40.0 0.6 35.8 m10 1,001 3 1,000 4.0630 0.0090 4.8253 0.0095 40.0 0.6 35.6 0 sh 1,000 4 1,000 4.0781 0.0073 4.8876 0.0099 40.0 0.7 36.0 p10 1,000 5 1,000 4.0077 0.0083 5.0359 0.0086 40.0 0.6 36.7 p20 1,000 4 1,000 3.9612 0.0096 5.0378 0.0088 40.0 0.5 37.4 p30 1,000 4 1,000 4.0405 0.0091 4.9798 0.0097 40.0 0.6 35.7 p40 998 6 1,000 4.0040 0.0079 5.0083 0.0082 40.0 0.5 36.6 p50 1,001 5 1,001 3.9707 0.0106 4.9803 0.0104 40.0 0.7 35.4 0.1 MPa LS (rel) case shaft locus was displaced with an overall averaged 33 \u00b1 4 \u00b5m variation, at proximity probe 1\u2019s location, and 3 \u00b1 4 \u00b5m variation, at proximity probe 2\u2019s location, as seen in Table 7. There was a subtle decrease in the standard deviation for the 0.2 MPa LS (rel) case (Table 7). The proximity probes measured an overall average variation of 41 \u00b1 4 \u00b5m for probe 1 and 4 \u00b1 4 \u00b5m for probe 2 for the shaft locus displacement with respect to the high-load reference tests (Table 7). Although the standard deviation of the load for the reference tests was greater than of the lower load tests, the proximity probes data were constant; hence, reducing the variation was not a main concern. Comparing the test results with previously mentioned experimental studies on pressure measurements, the study confirms the subatmospheric values, reported by other authors (Floberg (10), (11); Roberts and Hinton (12); Ikeuchi, et al. (14); Someya (15)). Some authors (Roberts and Hinton (12); Someya (15)) reported even lower than zero absolute pressures; that is, negative absolute pressures. Braun and Hendricks (9) provided more comprehensive pressure measurements in the axial direction than the current tests, but in the circumferential direction, their measurements were scarce compared to current tests. Furthermore, their tests were conducted on a much smaller transparent fully flooded bearing and hence used different supply geometry. Nevertheless, they were the first authors to publish 3D numerical reconstructions of the full pressure and temperature profiles obtained experimentally. Someya (15) obtained the highest negative absolute pressure peaks at the highest load tests. It is unclear, though, how the bearing sustained negative absolute pressures of 1 MPa, because the CGJB was open to the environment, and oil containing air cannot sustain tensile stresses. Mid-plane measurements are given in Someya (15). Roberts and Hinton\u2019s (12) tests revealed that negative absolute pressures are more likely to occur at coupled superlaminar and high eccentricity operation or for very high values of Reynolds number, Re \u2248 11,800\u201344,600. However, operating with a high clearance, the bearing should entrain a lot of air, which should not permit negative absolute pressures to develop. The average Reynolds number of the current tests, Re \u2248 0.0134, was significantly lower than the lowest Reynolds number, Re \u2248 43. In addition to providing experimental pressure field measurements D ow nl oa de d by [ U ni ve rs ity o f G ue lp h] a t 1 4: 20 0 2 O ct ob er 2 01 2 TABLE 7\u2014REFERENCE TEST VS. TEST CASES\u2014INFORMATION ON SHAFT LOCUS 0.1 MPa LS (rel) 0.2 MPa LS (rel) Delta (RC-TC) Proximity Probe 1 Delta (RC-TC) Proximity Probe 2 Delta (RC-TC) Proximity Probe 1 Delta (RC-TC) Proximity Probe 2 Shift Ave. (V) Ave. (\u00b5m) Ave. (V) Ave. (\u00b5m) Ave. (V) Ave. (\u00b5m) Ave. (V) Ave. (\u00b5m) m60 0.2671 35.5 0.0490 6.3 0.3333 44.3 0.0622 8.0 m50 0.3013 40.1 0.0165 2.1 0.3575 47.6 0.0269 3.5 m40 0.2688 35.8 0.0179 2.3 0.3261 43.4 0.0316 4.1 m30 0.2293 30.5 0.0536 6.9 0.2931 39.0 0.0600 7.8 m20 0.2517 33.5 0.0604 7.8 0.2984 39.7 0.0738 9.5 m10 0.2434 32.4 0.0760 9.8 0.3120 41.5 0.0833 10.8 0 sh 0.2179 29.0 0.0304 3.9 0.2755 36.7 0.0439 5.7 p10 0.2062 27.4 \u22120.0309 \u22124.0 0.2639 35.1 \u22120.0226 \u22122.9 p20 0.2834 37.7 \u22120.0217 \u22122.8 0.3357 44.7 \u22120.0119 \u22121.5 p30 0.2410 32.1 0.0005 0.1 0.2986 39.7 0.0125 1.6 p40 0.2410 32.1 \u22120.0190 \u22122.5 0.3053 40.6 \u22120.0033 \u22120.4 p50 0.2259 30.0 0.0305 3.9 0.3044 40.5 0.0331 4.3 at a usual clearance, the current tests were conducted with an increased measurements density, compared to the overall bearing dimensions. Regarding temperature distributions, the contour plots present an axisymmetrical characteristic, unlike Clayton and Wilkie\u2019s (20) asymmetrical distributions for a cantilevered endshaft. With respect to the extensive measurements performed by Mitsui, et al. (22), (23), temperature was measured at the fluidfilm bushing interface across four axial planes instead of three. Though the ratio of installed thermocouples on the bearing between the current tests and those of Mitsui, et al. (22), (23) was one order of magnitude lower, the number of locations for temperature measurements was higher. Similar to Clayton and Wilkie (20) and Mitsui, et al. (22), (23), far from the supply holes, the lightly loaded CGJB\u2019s axial gradients were moderate compared to the circumferential ones. Similar to Mitsui, et al. (22), (23), the temperature distribution revealed an axisymmetrical profile with respect to the land\u2019s mid-plane." + ] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure8-1.png", + "caption": "Fig. 8. Rear suspension of the small family car.", + "texts": [ + " 6, is used to present the developments reported in this paper and to serve as the object to carry the sensitivity analysis on the effect of the bushing properties on the vehicle dynamics. The data required to build the multibody model is obtained through direct measurement of the real vehicle components and, therefore, the manufacturer has no responsibility on the data presented or used here. The vehicle front suspension is a McPherson mechanism, presented Fig. 7. In the rear suspension, presented in Fig. 8, a torsion beam suspension system is used, which is a common choice for this vehicle segment as it insures the compactness required for a small car while reducing the need for an anti-roll bar. However, the vehicle considered here still includes the anti-roll bar. The inertia properties of the rigid body, center of mass location and body fixed frame orientations, and all data for the suspensions, tires and stabilization bars, are described in the work by Verissimo [9]. Two models of the vehicle are considered in the study that follows: one with ideal kinematic joints in the suspen- sion systems and another using bushing joint in selected suspension elements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003335_bf02126940-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003335_bf02126940-Figure2-1.png", + "caption": "Fig. 2.", + "texts": [ + " The resultant forces in every section i are obtained through the well-known expressions: X -M' o) h - 1 h't a . N ' + . N ' (2) of hg h = l In every section the constraints are (Fig. 1): gf,(M,, N t ) ~ a,~/I, + d f e N , - - &, <<. 0; i = 1 , 2, . . ., m; k = l , 2 , . . . , q (3) .. g i ~ g i t = ~ 9is =o Fig. 1. M; gi~ =0 Substituting (1) and (2) into (3) and stating that a must be maximum, we obtain the following linear programming problem (with unrestricted variables): a( , , , . ~ ; , + a ~ . ~ , ) + ~ x ~ ( , , , .~,~, + ~ , .~,~,) or else by (fig. 2, point Pz at a vertex) El = fll,'~ \" d|,'/7 + flt,~+l \" di,'~+! (11') (12') - - w <~ &~-- (a ,k \u2022 Moi + &e \u2022 NoO; i = 1, 2 . . . . . m; k = 1, 2 . . . . . q (4) 2 = max (5) 3. T h e d u a l p r o g r a m . The dual of the program (4), (5) is: Y X (a,,~\u00a2~, + d,,N'o,)u,, = 1 X ( ~ , ~ , + d , ~ , ) u , , = 0; h = 1, 2 . . . . . m y . b , ~ u , , - - X Y. (a,~\u00a2o, + d, dvo,)u,~ = rain (6) (7) (8) u,~> 0; i = 1 , 2 . . . . . m; k = l , 2 . . . . . q. (9) Moreover the solutions of the two programs in duality are to be chosen among those which satisfy the further conditions [8], [9I: (a,~M~ + d ~ - - b,~) \u2022 ,t~ = 0 00 ) i = 1 . . . . ,m; k = l . . . . , q 4. Interpretation of t h e d u a l program. Le~ us consider any mechanism with a maximum of ( 2 m - - r ) degrees of freedom which is obtained by placing the stress point on the yield curve in every calculus section of the structure. According to the normality law, the co~responding plastic strain components are given by (Fig. 2, point P~) with m, m.i., p,.~+x arbitrary positive coefficients, in number equal to or less than 2m. In order to obtain, with the Ot, ~, given by (11), (12) or (11)', (12)', the above-mentioned mechanism, we must satisfy some compatibility equations, r in number. As in [1], these conditions can be obtained by applying the virtual work principle to the mechanism itself, considering as systems of forces those systems constituted by the single redundant forces set equal to 1 and by the internal forces which equilibrate them in the primary system, that is: (0,#/~, + ~,N~,) = 0 , b = 1 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003126_1350650111424237-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003126_1350650111424237-Figure3-1.png", + "caption": "Fig. 3 Geometry of (a) intact and worn journal bearing and (b) misaligned journal bearing", + "texts": [ + " Equilibrium is attained when the sum of the calculated hydrodynamic forces and that of the externally applied loads are approximately equal (within a prescribed tolerance less than 10 4 of the load W ). The geometry of the worn bearing surface used in this analysis is the well-established known model presented by Dufrane et al. [8]. The total film thickness is the superposition of the film thickness corresponding to the misaligned journal bearing given by Nikolakopoulos and Papadopoulos [18] and the thickness h due to the abrasive wear of the bearing given by Dufrane et al. [8] depicted in Fig. 3(a) h\u00f0 , z\u00de \u00bc c \u00fe e0 cos \u00fe z y cos\u00f0 \u00fe 0\u00de \u00fe x sin\u00f0 \u00fe 0\u00de \u00fe h \u00f04\u00de where h \u00bc c\u00f0 0 1 cos \u00de \u00f05\u00de Equation (5) describes the change in the film thickness due to the bearing wear [35], and is applicable for the angles relevant to the worn region, otherwise h\u00bc 0. The worn zone is assumed to be centred on the vertical load direction, where 0\u00bcd0/c is the dimensionless wear depth (which corresponds to the percentage of the wear depth in relation to the radial clearance) and d0 the maximum wear depth. The worn zone is supposed to be centred to the ver- tical load direction and estimated [8,9] by the equation cos \u00bc 0 1 \u00f06\u00de where is the angle of the worn area" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure10-1.png", + "caption": "Fig. 10. Helical body with a helical radius of a and a pitch length of 2\u03c0b", + "texts": [ + " 9(b), the length of the body for one pitch is given by b0 = 2\u03c0r tan \u03b1. (1) Length of the wire for one pitch is w0 = 2\u03c0r cos \u03b1 . (2) When the wire is contracted, the body shape changes from a cylindrical to a helical. If the wire is contracted at a ratio of Kw, its length decreases as w1 = (1\u2212Kw) 2\u03c0r cos \u03b1 . (3) At the same time, the length of the body decreases as b1 = (1\u2212Kb)2\u03c0r tan\u03b1. (4) where Kb is the contraction ratio of the body, which depends on its elasticity. Let a and 2\u03c0b be the radius and pitch of the helical body, respectively, as shown in Fig. 10. Let r be the radius of the body. This helix can be expressed with a parameter \u03b8 as\u23a7\u23a8 \u23a9 x = a cos \u03b8 y = a sin \u03b8 z = b \u03b8 (5) Therefore, the length of the body for one pitch is given by b1 = \u222b 2\u03c0 0 \u221a( dx d\u03b8 )2 + ( dy d\u03b8 )2 + ( dz d\u03b8 )2 d\u03b8 = 2\u03c0 \u221a a2 + b2. (6) Length of the wire for one pitch is equal to the length of a helical line with a diameter of (a \u2212 r) and a pitch length of 2\u03c0b, because the contracted wire passes through the innermost route on the surface of the helical body. Therefore, the length of the wire is given by w1 = 2\u03c0 \u221a (a\u2212 r)2 + b2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001514_tmag.2009.2018676-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001514_tmag.2009.2018676-Figure6-1.png", + "caption": "Fig. 6. Simulation as using variable alpha angle at beta angle 0 degree.", + "texts": [ + " Because center positions of upper and lower coils are fixed at alpha 18 degrees, respectively, the stability of holding torque is simulated within alpha 18 degrees. In this simulation, the rotor rotates around the beta angle 30 degrees. Thus a magnet is affected from summed MMF created by 4 coils, for instance upper A, upper B, lower A, and lower B coil. Fig. 5 shows simulation results of rotor position at beta angle 30 degrees. In Fig. 5, a solid line is commanded position and a dot line is rotor position. The positioning error is gray area and its value is small. Thus, a rotor can tilt at commanded position well and position stably. Simulation of Fig. 6 shows different results. This simulation has variable alpha angle at fixed beta 0 angle. The rotor rotates around the beta angle 0 degree. Therefore a magnet is affected from summed MMF created by two coils, for instance upper A and lower A coil. Dissimilarly simulation result in Fig. 5, it shows that a rotor do not tilt to commanded position. Then there appears much position error. When there appears position error, a rotor cannot rotate stably. Figs. 8 and 9 are experimentation results. In Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure6-1.png", + "caption": "Fig. 6 Configuration 3 of the example", + "texts": [ + "url=/data/journals/jmdedb/27877/ on 02/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use J Downloaded Fr A2 = 0 1 0 0 1 1 0 \u2212 1 0 0 0 \u2212 1 0 1 0 0 0 1 0 1 1 0 0 1 0 9 It is interesting to know that if this configuration comes from Configuration 1, the adjacency matrix A2 can be obtained by multiplying \u22121 on A1 1,2 , A1 2,1 , A1 2,3 , and A1 3,2 . Moreover, according to Eqs. 5 \u2013 7 , n=n0 \u2212n\u22121=5\u22121=4, p=n1=4, and F=3 n\u22121 \u22122p=3 4\u22121 \u22122 4 =1. 3 If Link 3 is attached together with Link 4 by a pin, P, as shown in Fig. 6 a , the revolute pair between Link 3 and Link 4 is then frozen. The mechanism is termed as Configuration 3 and is shown in Fig. 6 b . If this configuration comes from Configuration 2, the adjacency matrix A3 can be obtained by multiplying \u22121 on A2 2,3 , A2 3,2 , A2 3,4 , and A2 4,3 , i.e., A3 = 0 1 0 0 1 1 0 1 0 0 0 1 0 \u2212 1 0 0 0 \u2212 1 0 1 1 0 0 1 0 10 4 If Link 4 is attached to Link 5 by a pin, P, as shown in Fig. 7 a , the revolute pair between Links 4 and 5 is then frozen. The mechanism is termed as Configuration 4 and is shown in Fig. 7 b . If this configuration comes from Configuration 3, the adjacency matrix A4 can be obtained by multiplying \u22121 on A3 3,4 , A3 4,3 , A3 4,5 , and A3 5,4 , i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure13.7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure13.7-1.png", + "caption": "Fig. 13.7 A closeup view of the sonar and infrared", + "texts": [ + " This resolution problem is magnified for objects further away from the robot (i.e., objects appearing at the wide end of the beam). Lastly, our robot is also equipped with an array of 11 Sharp GP20A21YK infrared proximity sensors above the sonar ring. The sonar and infrared sensors were mounted together so that their beams are 30913 Web-Based Control of Mobile Manipulation Platforms via Sensor Fusion BookID 175907_ChapID 13_Proof# 1 - 12/07/2009 oriented in the same direction. The configuration of sonar and infrared sensors is shown in Fig. 13.7. These sensors allow the RISCbot to obtain a set of observations to provide these observations to the controller and higher decision making mechanisms. The controller acts upon this set of observations to cause the robot to turn in the correct direction. The Integration of these modules together constitutes an intelligent mobile robot. A main drawback of the infrared sensors is that they can only accurately measure obstacle distances within a range of 0.1-0.8 m. Another drawback of these sensors is that they are susceptible to inaccuracies due to outdoor light interference as well as an obstacle\u2019s color or reflectivity characteristics which can be seriously affected by windows and metallic surfaces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003759_msec2015-9396-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003759_msec2015-9396-Figure1-1.png", + "caption": "Figure 1. Illustration of the principle of fused deposition modeling (FDM)", + "texts": [ + "org/about-asme/terms-of-use Technologies Example companies Materials Markets Fused deposition modeling (FDM) Stratasys (US) MakerBot (US) Bits from Bytes (UK)-now 3D Systems (US) Thermoplastic polymers Prototyping, casting patterns, end-use products Binder Jetting 3D Systems (US) Z-Crop (US) Voxeljet (Germany) Polymers, metals, foundry sand, ceramic Prototyping, casting molds, end-use products, direct parts Selective laser sintering (SLS) 3D Systems (US) Optomec (US) EOS (Germany) Polymers, metals, ceramic Prototyping, direct parts Stereolithography (SLA) 3D Systems (US) Formlabs (US) Envisiontec (Germany) Photopolymers Prototyping, casting patterns, jigs and fixtures including example companies, materials utilized, and typical markets are summarized in Table 1. 1.2 Fused deposition modeling (FDM) and filament extrusion With a current market share of 44%, thermoplastic-based additive manufacturing such as fused deposition modeling (FDM) is a prevailing technology [13]. FDM was first developed in 1989 by the Stratasys Inc., Eden Prairie, MN, USA [1]. As shown in Figure 1, thermoplastic filament is fed into the extrusion nozzle by the two filament drive wheels then extruded through the extrusion nozzle that trace the object model\u201fs cross sectional geometry. The resistance heater keeps the thermoplastic filament at a temperature just above its melting point. The thermoplastics harden immediately after flowing through the extrusion nozzle and bond to the layer below. After one layer is complete, the build platform lowers slightly to make way for the next layer. In the same principle, each single layer is built on atop the other until the object model is complete" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000819_aim.2009.5229935-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000819_aim.2009.5229935-Figure2-1.png", + "caption": "Fig. 2. Experimental setup.", + "texts": [ + " Under the background above, this paper presents a novel precise modeling methodology for the angular transmission error to improve the static positioning accuracy [10], where the nonlinear elastic component in micro-displacement region is mathematically modeled by applying a modeling framework for the rolling friction with hysteresis attributes [11], [12], as well as the conventional modeling for the synchronous component by spectrum analyses for rotation angle [4]. This transmission error model is then adopted to the positioning system as a model-based feedforward compensation manner, in order to improve the settling accuracy. The proposed modeling and compensation have been verified by a series of numerical simulations and experiments using a laboratory prototype. Fig.2 shows a schematic configuration of the laboratory prototype as an experimental positioning device, which is comprised of an actuator (AC motor) with an encoder, a harmonic drive gearing, an inertial load, and a load side encoder. Specifications of the prototype are listed in Table I. This positioning device is controlled in a typical semi-closed control manner by an angular feedback with an encoder mounted on the motor shaft, while the load side encoder measures and evaluates the load angle, i.e" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure2-1.png", + "caption": "Fig. 2. Radial misalignment in the cylindrical, revolute and translation joints.", + "texts": [ + " 1 the time derivative of \u03b4 is T \u03b4 \u03b4 = d d&& (5) where d& is the time derivative of Eq. (4) as P P j j j j i i i i= + \u2212 \u2212d r A \u03c9 s r A \u03c9 s& & % & % (6) where \u03c9 is the angular velocity of body i. When the bushing joint is not located in the center of mass of the connected bodies the transport moments are ( ) ( ) ' ' T P i i i i T P j j j j = = n A s f n A s f % % (7) where P is% is a skew-symmetric matrix made with the components of vector P is . The degrees of freedom to be penalized by this joint are the normal translational displacement, dn, depicted in Fig. 2, and the angular displacement due to the misalignment of the vectors, presented in Fig. 3. Based on Fig. 2 the misalignment of the axis of the joint in the tangential direction is given as ( )( )T T t j j j=d d s A s (8) while the misalignment normal to the joint axis in body j is n t= \u2212d d d (9) The unit vector si defines the joint axis in body i and the unit vector sj defines the joint axis in body j. The forces due to the normal misalignment of the axis are 1[ ( ) ( ) ]i n n n n n n j i K f b d \u03b4 \u03b4 \u03b4 \u03b4= \u2206 + + = \u2212 f d f f & (10) In Eq. (10) the magnitude of the deformation T n n n\u03b4 = d d and the velocities n\u03b4& and t\u03b4& are T n n n n\u03b4 \u03b4= d d& & (11) T t t t t\u03b4 \u03b4= d d& & (12) n t= \u2212d d d& & & (13) ( )( )T T t j j j=d d s A s& & (14) The moment due to the angular misalignment of vectors si and sj requires calculating angle \u03b8 between them", + " The formulation proposed for the cylindrical bushing joint requires the knowledge of the joint position, Pi and Pj, the coordinates of points Qi and Qj, which defines the joint axis, defined in bodies i and j, respectively. The translational stiffness defined in the normal direction, ( )nK \u03b4 , and the rotation stiffness ( )K \u03b8 , and their correspondent damping factors complete the full definition of the cylindrical bushing model input data. The formulation for the bushing revolution joint is based on the bushing cylindrical joint, to which a penalization of the relative displacement in the direc- tion of the joint axis is added. Based Fig. 2 this force is [ ( ) ( ) ] t i t t t t t j i K f b d \u03b4 \u03b4 \u03b4 \u03b4= \u2206 + + = \u2212 df f f & (20) The input for the revolute bushing joint is the same as the cylindrical joint, plus the stiffness defined in the tangential direction ( )tK \u03b4 and the corresponding damping coefficient b. The models of the bushing joints require that their stiffness is defined. Because the bushings are rubber type materials, their stiffness is nonlinear and characterized by functions that need to be identified. For the purpose, four test cases were conducted in the finite element (FE) program ABAQUS [8] with an FE model of the bushing element" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000947_tmag.2009.2012590-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000947_tmag.2009.2012590-Figure5-1.png", + "caption": "Fig. 5. Field distribution of the 6/4 SRM for two phase excitation and 40 rotor angular displacement from the aligned position. The current of phase A is 10 A and of phase B is 5 A.", + "texts": [ + " As primary active phase, we choose the phase with the maximum observed flux linkage, whereas as secondary active phase the one with the intermediate flux linkage level. By examining the estimation results produced by the commonly used method of linear interpolation based only on the self-flux linkage data, one can perceive that the estimation errors that arise are significant, especially around the aligned position for the primary active phase. These errors are mainly imputed to two factors: the magnetic saturation and the interaction between the two active phases. In Fig. 5, the resultant field distribution of a 6/4 SRM for two phase excitation is depicted. The resultant flux linkages from this field distribution can be interpreted as a suitable superposition of the existing data of the self- and mutual flux linkages for single phase excitation, as shown in Fig. 6. Thus, the flux linkage of the phase that is positioned at zero degrees in Fig. 6 includes two contributions: one that can be attributed to the self linkage of the primary active phase and one that can be attributed to the mutual linkage of the secondary active phase" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003050_978-3-642-36279-8_15-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003050_978-3-642-36279-8_15-Figure9-1.png", + "caption": "Fig. 9 Exploration and coverage of an office environment by a team of 4 robots. Blue curves indicate boundaries of tessellations, intensity of white indicates the value of entropy.", + "texts": [ + " In addition, to avoid situations where a robot gets stuck at a local minima inside its tessellation even when there are vertices with entropy greater than \u03c4 in the tessellations (this can happen when there are multiple high entropy regions in the tessellations that exert equal and opposite pull on the robot so that the net velocity becomes zero), we perform a check on the value of the integral of the weight function, \u03c6 , within the tessellation of the kth robot when its control velocity vanishes. If the integral is above the value of \u222b Vk \u03b5\u03c6 dq, we switch to a greedy exploration mode where the kth robot essential head directly towards the closest point that has entropy greater than the value of \u03c4 . This ensures exploration of the entire environment (i.e. the entropy value for every accessible vertex drops below \u03c4). And once that is achieved, both \u03c6 and \u03b6 become independent of time. Thus convergence is guaranteed. Figure 9 shows screenshots of a team of 4 robots exploring a part of the 4th floor of the Levine building at the University of Pennsylvania. The intensity of white represents the value of entropy. Thus in Figure 9(a) the robots start of with absolutely no knowledge of the environment, explore the environment, and finally converge to a configuration attaining good coverage (Figure 9(d)). Figure 10 shows a similar scenario. However, in this case one of the robots (Robot 0, marked by red circle) gets hijacked and manually controlled by a human user soon after they start cooperative exploration of the environment. That robot is forced to stay inside the larger room at the bottom of the environment. Moreover, in this case we use a team of heterogeneous robots (robots with different sensor footprint radii), thus requiring to compute power voronoi tessellations. This simple example illustrates the flexibility of our framework with respect to human-robot interaction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001076_s1068366610040100-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001076_s1068366610040100-Figure2-1.png", + "caption": "Fig. 2. Phase plane portraits of system at values of parameters \u03b8 = 0.7, g\u03b8 = 0.4, \u03b1 = 0.1, \u03b2 = 4, and Te = 25 and at noise intensity D = 0, which represent (a) solution of equation (9) at \u03c4 = \u03c4\u03b5/\u03c4\u03c3 = 15; (b) solution of equation (15) at \u03c4 = \u03c4T/\u03c4\u03c3 = 120; and c\u2014solution of equation (16) at \u03c4 = \u03c4T/\u03c4\u03b5 = 120.", + "texts": [ + " \u03be t( )\u2329 \u232a 0; \u03be t( )\u03be t '( )\u2329 \u232a 2D\u03b4 t t '\u2013( ),= = \u03c3\u00b7\u00b7 2\u03b3\u03c3\u00b7 \u03c90 2\u03c3+ + \u03be t( ),= (\u03c3\u00b7 To solve equation (11) numerically, we make the substitution y = yielding (12) where \u03b3 and \u03c90 follow from the comparison of (9) and (11). Then we use the Euler method for integration. In this case, the iteration procedure is as follows [11, 12]: (13) We use the Box\u2013Muller model [19] to simulate the random force Wn: (14) where the pseudorandom numbers r1 and r2 are uni formly distributed. The numerical solution of equation (9) obtained using procedure (13), (14) is presented in Fig. 2a in the form of a phase plane portrait. Here, the isocline along which = 0 is shown by the dashed line and the phase trajectories have the vertical tangent. This iso cline is the abscissa axis of the coordinates under con sideration. The dotted line represents the isocline = 0, along which the phase trajectories have the horizontal tangent. Since the attenuation factor in (9) depends on when = 0 equals zero, the expression for the isocline (\u03c3) is found from the solution of the quadratic equation; this is why the dependence is binary", + " For the considered ratio between the relaxation times, such behavior appears within the whole range of the parameters [17]. We note that positive and negative values of the stresses corre spond to the motion of the upper friction surface in different directions. Thus, at a positive initial value of \u03c3, which is proportional to the shear rate, and a nega tive initial value of or acceleration, reciprocal motion may occur, as follows from the figure. Figure 3a illustrates the time dependences of the stresses, which correspond to the trajectories shown in Fig. 2a. The dependences present the nonperiodic transient mode, when values of the stresses vary until a constant sliding velocity is reached (\u03c3 = const). Figure 4a shows the solution of the same equation as Fig. 3a but at D \u2260 0. It is seen that with time the stresses vary randomly, yet within a narrow range since the noise intensity is low; this corresponds to the sliding mode. The dependence is presented starting from the moment t = 1000 since in this study we consider the stationary friction mode rather than the transient one", + " The dependence S(\u03bd) contains no pronounced maxima, which proves the absence of the periodical component in the dependence \u03c3(t). Thus, in this case the sliding mode with a slightly fluctuating shear rate becomes more steady with time. The case \u03c4\u03b5 \u03c4T, \u03c4\u03c3. In this case, we use the approximation \u03c4\u03b5 \u2248 0 in the original system; if the time is measured in units of \u03c4\u03c3, this yields the following equation: (15) where the ratio \u03c4 = \u03c4T/\u03c4\u03c3 is introduced and the coeffi cients A and B are determined in (9). The phase plane portrait resulting from the solu tion of (15) is shown in Fig. 2b. It is seen that in this figure the same singular points appear, but the differ ence is that the points O and O ' transform to stable focuses and damped oscillations occur in the system. The figures indicate the phase trajectories. The iso \u03b5\u00b7 \u03c3\u00b7\u00b7 A \u03c3 1\u2013\u2013( )\u03c3\u00b7 A\u03c3 \u03c4 1\u2013 1 \u03c32+( )+ +[ ]\u03c3\u00b7 B\u03c3+ + = \u03be t( ), JOURNAL OF FRICTION AND WEAR Vol. 31 No. 4 2010 A STOCHASTIC MODEL OF STICK SLIP BOUNDARY FRICTION 313 cline shown by the dotted line differs from the isocline in the previous case. The corresponding time dependences (Fig", + " Moreover, the effect of fluctuations may cause the instability of the focus, resulting in continuous increase in the ampli tude of stress oscillations, which is similar to reso nance in the system. The case \u03c4\u03c3 \u03c4\u03b5, \u03c4T. In this case we assume \u03c4\u03c3 \u2248 0 and measure time in units of \u03c4\u03b5, yielding (16) \u03c3\u00b7 \u03c3\u00b7\u00b7 2A2\u03c3 A\u03b2 2A\u03b2\u03c3\u03b2 \u03b1\u03b2 \u03c3\u03b2+ \u2013 1 \u03c3 \u2013+\u239d \u23a0 \u239b \u239e \u03c3\u00b7 A\u03c3 1 1 \u03c4 +\u239d \u23a0 \u239b \u239e 1 \u03c4 + ++ \u00d7 \u03c3\u00b7 A\u03c3 1+ B A\u03c3 1+ \u03c3+ \u03be t( ),= 314 JOURNAL OF FRICTION AND WEAR Vol. 31 No. 4 2010 KHOMENKO, LYASHENKO where \u03c4 = \u03c4T/\u03c4\u03b5. At this ratio between the relaxation times, before the stationary state is reached, a greater number of oscillations occur around the focus on the phase plane portrait (Fig. 2c) compared to the previous case. This is also confirmed by Fig. 3c, which illustrates long oscillations that do not attenuate even at t = 900. Figure 4c shows the time dependence of the stresses under the effect of noise (in all of the cases under con sideration D = const); it is flatter and more regular than that presented in Fig. 4b. The corresponding spectrum (Fig. 5c) has a much narrower and higher peak at \u03bd \u2248 0.06 than the spectrum shown in Fig. 5b; this is why the fundamental frequency is \u03bd \u2248 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000233_tie.2007.898297-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000233_tie.2007.898297-Figure1-1.png", + "caption": "Fig. 1. MEMS/IC assembly approach.", + "texts": [ + "00 \u00a9 2007 IEEE investigation [14]. However, recent progress in the field is promising enough to consider mixed technology using hybrid and flip-chip integration, as it can be applied for multichipmodule (MCM) technique, which has been successfully developed using low-cost IC solutions as application-specified ICs (ASICs) [15], [16]. The authors are interested in combining a low-cost chip, the field-programmable gate array (FPGA) with MEMS chip, and assembling them using MCM technology, as shown in Fig. 1. In an FPGA component, the hardware can be reconfigured. FPGAs emerged as a new technology for the implementation of digital logic circuits during the mid-1980s. The basic architecture of an FPGA consists of a large number of configurable logic elements and a programmable mesh of interconnections. New architectures of FPGAs both lower the cost and improve functionality, enabling them to cost-effectively compete with ASICs (even for high fabrication volume density). This technology has been used for motion control [17], [18], and is ready to be applied to new applications as MEMS, particularly for embedded and distributed devices [19]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001756_la803390x-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001756_la803390x-Figure5-1.png", + "caption": "Figure 5. Images for the fibrillar sample with w ) 80 \u00b5m sliding at 5 \u00b5m/s under a SAM-coated cylindrical indenter with a 1 mm radius. (a) Initial contact and (b,c) two consecutive still frames illustrate the sudden recovery of one fibril located near the leading front (see inset circle).", + "texts": [ + " However, the sliding friction remained unaffected by this damage since sliding friction forces were considerably smaller and were measured over an undamaged region of the sample. Figure 4 also shows that the sliding friction is much lower than the static friction and is practically independent of spacing, as observed with a spherical indenter.4 Using a synchronized video, pictures of the contact zone of all fibrillar samples were taken with a frequency of 1.5 Hz. To illustrate how we measure deformation of fibrils in the contact zone, we present several snap shots of a sample with w) 80 \u00b5m. Figure 5a is the first still frame of the contact region. The top ends of the fibrils appear as fuzzy gray circles since the joint between the fibrils and the film is rounded (see Figure 1a). The small darker squares indicate the joints between the bottom of fibrils and the thick backing layer. In Figure 5a, no shear has been applied, so the squares and circles overlap. When sheared, they are mutually offset (see Figure 5b,c). By measuring the distance between them, one can determine the relative deflection of each fibril. To capture the micromechanics of fibril deformation before sliding, we analyzed a sequence of snap shots for a microfibril array with w)65 \u00b5m to determine the deflections of one particular row of fibrils as a function of time. Since the focus of this paper is on sliding friction, results of deformation before sliding initiates are shown in the Appendix 1. The salient new finding is that overall sliding is preceded by some local microslip near the trailing edge. Figure 5b,c is from two consecutive frames capturing a sudden partial recovery of shear in one fibril (highlighted by a circle) near the leading edge of the contact during sliding. Because the sample is moved at a constant rate, the deformation history of each fibril entering the contact zone should be approximately the same. To study this deformation history, we marked a single row of fibrils aligned with the direction of sliding (highlighted in Figure 5b,c) with symbols 3-8. The deformation histories of these fibrils during sliding are followed by measuring the fibril deflections on a sequence of consecutive still frames. Figure 5b,c shows that the microscope view range is larger than the contact width, and the number of fibrils in a row is limited to 6. This means that the history of some fibrils can be traced only partially. Figure 6 plots the time history of deflection for fibrils 1-10. The time origin has been set to the time when fibril 1 just exits the contact zone. Note that fibrils 1-5 are already out of the leading edge, so complete deformation histories are not recorded. As each fibril enters the contact zone, it first shears at a rate determined by indenter velocity, shown by the slanted lines (fibrils 7 - 10). At a critical shear, a fibril suddenly slips backward (see also Figure 5b,c) and releases some of its elastic energy. This causes the previously buckled film between this fibril and the fibril to its right to unbuckle partially. These slip-stick events occur repeatedly with decreasing magnitude and are shown in more detail for a particular fibril in Figure 7. Note that, at the end of each slip/stick event, the fibril reloads at the approximately the same displacement rate of the indenter (5 \u00b5m/s), as shown by the slanted lines in Figure 7. As mentioned earlier, we expect that all the fibrils have a similar deformation history", + " Figure 8 shows a very important feature of our film-terminated microfibril array that is absent in an array without a continuous terminal film. It is that the shear force carried by the film outside the contact zone (behind the trailing edge) is nonzero and can actually be greater than the force carried by the fibrils inside the contact zone. The continuous film allows the shear force to be transmitted to noncontacting fibrils, as shown by the shear displacements of fibrils outside the contact zone in Figure 5b,c. The width of the noncontact region where shear force is nonnegligible is defined as the shear lag width. We have found that, with increasing spacing between fibrils, the shear deflection on each fibril and the shear lag width also increase, counteracting the decrease in fibril density and maintaining a constant total shear force.10 Two questions remain. Why is the sliding friction of our fibrillar sample independent of spacing (geometry) and displacement rate? In addition, why is the sliding friction for fibrillar samples so close to that of the flat control sample" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001497_j.matdes.2010.02.029-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001497_j.matdes.2010.02.029-Figure12-1.png", + "caption": "Fig. 12. Colour to gray scale conversion.", + "texts": [ + " 2D cross sections of crane hook in XY plane along with stress values have been stored as jpg image (Fig. 10). Further stress ranges can be chosen as per the requirements or number of regions one wishes to have. In the present case three stress ranges have been considered as shown in Fig. 11. A MATLAB program is developed to generate CLI file from jpg images stored through APDL, corresponding to each layer of the part. Coloured image is first converted into gray scale image and then contours are separated based on gray values (Fig. 12). Coordinates of pixels at boundary of each contour are written in the form of CLI file through developed MATLAB program for each subpart/ colour (Fig. 13). Procedure for separating contours of different stress levels and writing CLI file is described in Fig. 14. Further a MATLAB program is written separately to read and plot the contents of CLI file for visualisation and debugging purpose. Contours for three consecutive layers for all three colour1 (which represents stress values) ranges in this case are presented in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003545_mesa.2012.6275564-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003545_mesa.2012.6275564-Figure3-1.png", + "caption": "Figure 3. Influence of fluid state on pressure distribution", + "texts": [ + " As shown in graph: The calculation results are in good agreement with Gabriel\u2019s data, the gas film thickness has great influence on pressure and the influence is getting greater while the gas film thickness is reducing, then the effect of dynamic pressure is increasing (as shown in the figure (a)), meanwhile the pressure peak forms in the boundary of spiral groove and seal dam; Conversely, when gas film thickness is deeper, the gas film pressure will decrease, and the dynamic pressure effect is slighter even disappear (as shown in figure (c)). V. DISCUSSION OF RESULTS The calculation parameters agreement with the parameters which used in classic literature wrote by Gabriel R P (1994), the closed force is 33.1687 KN. A. Influence of Fluid State 1) Influence of Fluid State on Pressure Distribution: Figure 3 presents a view of influence of fluid state on pressure distribution. As shown in figure: (1) As the initial film thickness of S-DGS are gradually increasing from 2.03\u03bcm to 5.08\u03bcm (2.03\u03bcm 3.05\u03bcm 5.08\u03bcm), the pressure distribution gap between the laminar and turbulent is smaller and smaller; This is because along with the increasing of film thickness, the influence of viscous resistance between seal face on the gas flow is down, thereby reducing the influence of viscosity on pressure distribution; (2) The most obvious difference between with laminar model and turbulent model is turbulence model could be able to enhance dynamic pressure open characteristics and face pressure (especially the root of spiral groove), this is because when fluid is in laminar situation, the gas is ideal and free fluid, but using RNG k- turbulence model, medium gas will be seen as incompressible viscous fluid, and in fact, gas is sticky" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001741_j.commatsci.2009.02.007-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001741_j.commatsci.2009.02.007-Figure4-1.png", + "caption": "Fig. 4. Finite element models near brazed joints with various heights; (a) the zero height by multi point constraint (MPC), (b) 2 mm, (c) 4 mm, and (d) 6 mm heights.", + "texts": [ + " The unit cell was modeled by using a commercial graphics code, 3-D PATRAN 2005 as shown in Fig. 3. The finite element analysis was performed by using ABAQUS version 6.5. The cross section of the wire was modeled by 20 quadratic brick elements (C3D20 element of ABAQUS) and the filler metal brazed at the cross points (see the figure) was modeled by 5802 quadratic tetrahedron elements (C3D10 element of ABAQUS). A variety of different heights of the filler metal from 0 to 6 mm was used to investigate the size effect of the brazed part (see Fig. 4). The unit cell model for the analysis was composed of 145,992 elements and 331,174 nodes in total. The material properties for the simulation were given by the tensile test by Lee et al. [12]. The elastic modulus of the wire and the filler metals was E = 170 GPa, the yield stress was ry = 184 MPa, and the Poisson\u2019s ratio was assumed to be m = 0.3. The J2-incremental theory of plasticity was applied. l wires, in which the triangles surrounded by the solid line and the parallelogram parallelepiped in (d) the three-dimensional Kagome truss woven by six directional When a structure consisting of many uniform cells like WBK is to be analyzed, numerical analysis for the whole structural system can be severely inefficient or even impossible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002698_iros.2012.6385850-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002698_iros.2012.6385850-Figure4-1.png", + "caption": "Fig. 4. a) Intersection of the perpendicular bisector of (q0, qj) and the circle CR gives rise to the chord Lj and the half-plane induced by qj containing q0. b) The chord L2 is removed by CheckChords method, as it does not contribute to V R.", + "texts": [ + " Phase 1: Chord construction: The boundary of kV R is made up of line segments corresponding to perpendicular bisectors and arcs on the circle CR. First we look at the intersection of b0j , the perpendicular bisector of line joining q0 and qj (j-th node on Q, the relative configuration sorted based on radius) with the circle CR. Let the points of intersection of b0j with CR be (R, \u03b8sj ) and (R, \u03b8ej ). Since R is known, we need to find only \u03b8sj and \u03b8ej , which are given by: \u03b8sj = mod 2\u03c0(\u03b8j + \u03b4\u03b8j) \u03b8ej = mod 2\u03c0(\u03b8j \u2212 \u03b4\u03b8j) (2) where, \u03b4\u03b8j = cos\u22121 ( rj 2R ) . This is illustrated in Figure 4(a). Let Lj denote the segment of b0j that lies between (R, \u03b8sj ) and (R, \u03b8ej ). Note that Lj forms a chord within CR corresponding to qj . A segment of Lj \u2208 \u2202V R if and only if qj \u2208 NLD(P, 2R, pi), where \u2202V R represents the boundary of V R. Further note that, if pj \u2208 P 2R i \u2229ND(P, pi), then qj \u2208 NLD(P, 2R, pi). Starting with C1, the i-th robot computing V R i constructs the chords corresponding to robots in Q in successive virtual circles and stores them in a set denoted by L. The cumulative set of chords L at the k-th virtual circle Ck is the given by: L = L \u222aj\u2208{l|ql\u2208Ck} Lj ", + " It takes Q, the set of robots that are in P2R i as input and returns the angle-sorted set chords L corresponding to each qj \u2208 Q. It also constructs a sorted set \u0398 containing the polar angles of the endpoints of every chord in L, which is used in the next phase of the construction of the RCVC . Phase 2: Removing Redundant Chords: Let Lj , Lk \u2208 L denote two chords corresponding to points qj and qk respectively. If Lk lies entirely within the half-plane of Lj that does not contain pi, then Lk /\u2208 \u2202V R. As illustrated in Figure 4(b), L2(= Lk) is entirely within the half-plane of L1(= Lj) that does not contain pi. In such a case, Lk is constructChords(Q) Input: Q// Q the relative configuration of robots in P2R i in polar coordinates with pi as reference and sorted based on radii. Output: L,\u0398, // L: set of chords corresponding to nodes in Q, \u0398: set of endpoints of chords in L within CR L \u2190 \u2205; \u0398\u2190 \u2205; foreach qj \u2208 Q do Lj \u2190 Perp. bisector of line (q0, qj) : qj \u2208 Q; (\u03b8sj , \u03b8 e j )\u2190 Angles of extremities of chord Lj with q0 within circle C(q0, R) of radius R; L \u2190 L \u222a Lj ; \u0398\u2190 \u0398 \u222a {\u03b8sj , \u03b8 e j} end Sort L in ascending order of angle of corresponding nodes; Sort \u0398 in ascending order; return L,\u0398; Algorithm 1: Method for constructing chords for robots (nodes) that are within twice the sensor range of robot i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure4-1.png", + "caption": "Fig. 4 With the transmission ratios k1 and k2, the double pendulum can be balanced with only one CRCM", + "texts": [ + " The equations of the reduced inertias of this configuration can be calculated as before and become I 1 red = I2 + m2l2 2 + m2 l2 2 1 \u2212 k2 + m1 l1 2 + m2 + m2 l1 2 + k1 2I1 24 I 2 red = I2 + m2l2 2 + m2 l2 2 + k2 2I2 25 where Eq. 22 was substituted into Eq. 24 . Since the CRCM m2 in this configuration balances the moment of link 2 for any motion of the mechanism, the reduced inertia about O is lower than that of the basic CRCM configuration. 2.3 Balancing With Only One CRCM. Due to the appearance of transmission ratio k2 in Eq. 21 , there exists a situation for which I1 can be fixed to link 1, and the double pendulum is moment balanced with only one CRCM. Figure 4 shows a possible way in which, with respect to the configuration of Fig. 3, the gear at A is driven with a fixed gear at O and with transmission NOVEMBER 2009, Vol. 131 / 111003-3 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use r m i w s B a S f a g n a b F l m F c m t 1 Downloaded Fr atio k1. m1 in this case can be fixed to link 1 and is a counter- ass. The angular momentum of this mechanism writes hO,z = I2 \u03071 + \u03072 + I2 \u03072 + I1 \u03071 + e r2 m2r\u03072 + e r1 m1 r\u03071 + e r2 m2 r\u03072 26 n which the kinetic relation of \u03072 depends on k1 and k2 and is \u03072 = 1 \u2212 dO dA,1 dA,2 dC \u03071 + 1 \u2212 dA,2 dC \u03072 = k1 1 \u2212 k2 \u03071 + k2\u03072 27 ith dO, dA,1, dA,2, and dC being the diameters of the gear at O, the mall gear at A, the large gear at A, and the gear at C, respectively", + "5 Evaluation of the Different Configurations. The inertia of the CRCMs and the inertia equations depend on the choice of the transmission ratios. For increasing transmission ratios, the inertia of the CRCMs decreases proportionally, but the equations for the reduced inertia increase quickly since the transmission ratios appear squared. Practically, the advantage of the configuration of Fig. 3 is its low inertia since also for motion about O, the moment of link 2 is balanced by CRCM m2 . The advantage of the configuration of Fig. 4 is that only one CRCM is necessary for the moment balance of the complete mechanism. The configuration of Fig. 5 is useful since having CRCMs only near the base allows a compact construction of a balanced machine. In addition, both actuators can be mounted on the base compactly, by which the chain that transmits the motion of link 2 to the base for balancing purposes is also used to drive link 2. The inertia equations of the configuration of Fig. 4 can be re- duced by combining it with Fig. 3, where m1 in Fig. 3 is driven by Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use a o d d f t b r o l m 3 m p t q p p o n n p b n m s b 7 C n l n C s t d i s A t F n J Downloaded Fr parallel transmission k1=0 and does not rotate for any motion f the mechanism. m2 is then used for the moment balance of both egrees of freedom, as in Fig. 4. The result is that the term I1 rops from Eqs. 26 , 28 , 29 , 31 , and 32 . It is evident that or a low inertia, the countermasses should not rotate with respect o the base, besides what is necessary to maintain the moment alance. Limitations of the CRCM principle can be the transmission atios that have to remain relatively small in practice, a gear ratio f 8 is already high , and the shape of a CRCM that may become arge to obtain the specific amount of inertia and mass. These and ore limitations are studied in Ref" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001118_09544054jem1847-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001118_09544054jem1847-Figure1-1.png", + "caption": "Fig. 1 Dimensions of specimen", + "texts": [ + " Since the residual stress and microhardness distribution is a critical factor for fatigue performance, this study investigates the residual stress and microhardness distribution of a specimen without a white layer and one with a white layer. The fatigue parameters are computed based on the residual stress and microhardness distribution to investigate its effect on the fatigue performance. Fatigue tests are then performed to substantiate the computations. Specimens of AISI 52100 steel were prepared for the experiment, since this steel is widely used for hard machining and bearing applications. The composition of AISI 52100 steel is listed in Table 1. The dimensions of the specimen (Fig. 1) were selected to be uniformly through-hardened and to minimize a deflection by the chucking forces of a standard jaw [12]. The cutting tool used in the experiment was a cubic boron nitride (CBN) tool (BRNU-42 BZN 8100). Two different combinations of machining conditions were determined (Table 2). Condition A was selected to avoid a white layer at the machined subsurface, while condition B was selected to induce a thick white layer at the machined subsurface [13]. The specimens were machined by using a Fryer SL10 CNC lathe", + "6 Table 2 Machining conditions Parameter Condition A Condition B Cutting speed (m/s) 1.41 2.82 Feedrate (mm/s) 0.05 1.66 Depth of cut (mm) 0.1 0.2 Coolant Minimum quantity lubrication Dry Flank wear land (mm) 0 0.1 Proc. IMechE Vol. 224 Part B: J. Engineering Manufacture JEM1847 at Purdue University Libraries on June 2, 2015pib.sagepub.comDownloaded from X-ray diffraction was used to measure the residual stress distribution in the subsurface of specimens. Residual stresses were measured in the cutting direction (Fig. 1). A Denver-Proto XRD 3000 residual stress analyser was used with a CrKa radiation tube. The sin2c technique was applied to compute the residual stress value [15]. Nine c angles were used for the computation: 20, 15, 10, 5, 0, 5, 10, 15, and 20 . The measurements were taken at the exposed surface and ten different depths: 2.5, 5.1, 7.6, 10.2, 12.7, 25.4, 38.1, 50.8, 88.9, and 127mm. For the measurements at different depths, a particular amount of a layer was removed with an electrolytic etcher, saturated NaCl solution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001491_09544062jmes1452-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001491_09544062jmes1452-Figure3-1.png", + "caption": "Fig. 3 Design of the inlet and outlet ports", + "texts": [ + "8 By using the formula given above one can work out each angle values. For reaching the effective inlet and outlet ports during operation, the design of ports JMES1452 \u00a9 IMechE 2009 Proc. IMechE Vol. 223 Part C: J. Mechanical Engineering Science at BRIGHAM YOUNG UNIV on June 25, 2015pic.sagepub.comDownloaded from becomes important. First, the distribution angle of the inlet port should be bigger than the outlet port, which can make an internal pressure and increase the air-extracting ratio. If the design is shown in Fig. 3, the operation principle of intake and exhaust is shown in Fig. 4. First, the two rotors rotate in opposite directions. Figures 4(a) to (f) represent a cycle rotating. When the rotors rotate as given in Fig. 4(a), the inlet and outlet ports are sealed. Coming to Figs 4(b) and (c), the angle range of the inlet port and the capacity would increase. This is the extracting process and it is a low-pressure area. At the outlet port, the capacity may change from large to small, so the air would be pressed", + " Then it begins to exhaust while the capacity becomes small, the outlet port is a high-pressure area. The inlet and outlet ports would be sealed again till Fig. 4(e). In For a clearer concept of the design, Fig. 5 shows the rotor motion locus. Because the parameter \u00b5 is set 0.8, there is no mesh in the crown and the root of the rotor profile. If the ports of the inlet and outlet are not sealed at these positions (see Figs 4(e), (f), and (a)), the gas may leak out through these two ports. Hence, this would affect the pump performance. If the inlet and outlet design is as shown in Fig. 3, some high-pressure air is represented as the dark area in Fig. 4(e) may go through Figs 4(e) and (f) to Fig. 4(a), finally it would be brought to the inlet area. Under a fixed value \u03b4, \u00b5 is taken as 0.7, 0.8, 0.9, and 1. Here, \u03b4 is given as 1.5, 1.6, and 1.7. If r = 40, the rotor profile can be derived from the above equation. Figure 6 shows results for different \u00b5 (\u03b4 = 1.5). It is obvious from the figure that the distance is becoming far from one claw tip of the rotor to the other rotor profile when \u00b5 becomes small. However, the inlet and outlet ports design is given in Fig. 3. If the angle of the Proc. IMechE Vol. 223 Part C: J. Mechanical Engineering Science JMES1452 \u00a9 IMechE 2009 at BRIGHAM YOUNG UNIV on June 25, 2015pic.sagepub.comDownloaded from port is too large, then some air may be produced from the outlet port to the inlet port (see Fig. 4(e)). This may lead to a low air-extracting efficiency. If the port is like the design of Fig. 3, the carryover area can be defined as Ac, showed as Fig. 7. Besides, for evaluating the area efficiency of the rotor, this article may call it as an invalid area, which means when the inlet and outlet ports are sealed, it is useless for extracting the air no matter how much area is present between the two rotors. But, for a convenience to compare different parameters, this aricle would take the invalid area Anf as three different positions (see Fig. 7). The sum of the carryover area Ac and another two invalid areas Anf1 and Anf2, is represented as Anf = Ac + Anf1 + Anf2 (22) The definition of area efficiency is as \u03b7 = Ach \u2212 Anf \u2212 2Aro Ach (23) where Ach is the chamber area and Aro is the rotor area", + " However, the pump efficiency cannot just consider area efficiency. The clearance may also affect the pump efficiency. To prevent air leaking from the clearance, usually at the top of the rotor an arc whose Fig. 13 Design of the outlet port (\u00b5 = 1) radius is close to the chamber radius would be used. The wrap angle of the arc depends on the working condition. For a better pump performance, the inlet and outlet ports design are also important. For example, if the parameter is set as \u00b5 = 1 and the positions of the inlet and outlet ports are same as Fig. 3. When two rotors rotate to the position like Fig. 4(d), the inlet port is closed and the outlet port is open (high-pressure area) to exhaust gas until the position of Fig. 4(e). Some highpressure gas would be brought from Figs 4(e) to (f). However, at the position of Fig. 4(f) (same as Fig. 6, \u00b5 = 1), the high-pressure area would be divided into three sections that may produce large noise, and also make invalid work. To solve the problem, the size and the position of the outlet port can be designed as Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002472_1.3554974-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002472_1.3554974-Figure1-1.png", + "caption": "Fig . 1 Bearing geometry", + "texts": [ + "2 0.5 ^ \u2014 ^ 1 h, = 1.84C LYU \u2122\u00b0- 731 (\u00a3) D \u00a3 \\ 0 . \u00bb s (S) 0.43 0 . 5 6 / C 4 V ' \" At = -. . - P (S) l ^ D j (I) 2*P - 0.128 (|)\u00b0'048 \u00ab \" \" / ( * ) \" \" (S)\"\u2022/(\u00a3)' 2.1 2.2 S Z 0.15 L 0.25 ^ \u2014 ^ 0.5 fto = 1.035C(-^rJ (S) / V W ( r \\ 0.04* B ) JT m a x RNCL (S)\u00b0 2 . 9 ^( s and \u03b2/2 < \u03c0/4 (because \u03b1 > \u03c0/2), the following is valid: l2t+1 < l2t + \u03b52 \u2212 2lt\u03b5 \u221a 2/2 < l2t + \u03b52 \u2212\u221a 2lt\u03b5 < l2t + \u03b52 \u2212\u221a 2\u03b5s. (8) This allows to bound the decrease in the squared distance between robots in one iteration as l2t \u2212 l2t+1 > \u03b5( \u221a 2s \u2212 \u03b5). (9) 3) 0 moves, 1 stops: This situation is illustrated by Fig. 1. Here, tanker 0 moves to point x, decreasing the distance to the worker from lt = \u2016r0 \u2212 r1\u2016 to lt+1 = \u2016r1 \u2212 x\u2016. The distance tanker travels is \u03b5 = \u2016x \u2212 r0\u2016. By the law of cosines l2t+1 = l2t + \u03b52 \u2212 2\u03b5lt cos \u03b3. (10) To bound \u03b3, we apply the law of sines, w0/ sin \u03b3 = w1/ sin(\u03b1 \u2212 \u03b3). Starting with w0 sin(\u03b1 \u2212 \u03b3) = w1 sin \u03b3 and doing trigonometric transformations, we derive tan \u03b3 = w0 sin \u03b1 w1 + w0 cos \u03b1 . (11) Since \u03b2 \u2264 \u03c0 \u2212 \u03b1, cos \u03b2 \u2265 cos(\u03c0 \u2212 \u03b1). This inequality can be plugged into stopping condition (5) to get 2w2 0 (1 + cos(\u03c0 \u2212 \u03b1)) < w2 1 . The latter simplifies to cos \u03b1 > (2w2 0 \u2212 w2 1 )/2w2 0 . Plugging this into (11) and using w1 \u2264 2w0 , we obtain tan \u03b3 < w0 sin \u03b1 w1 + [(2w2 0 \u2212 w2 1 )/2w0 ] < 2w2 0 2w0w1+2w2 0\u2212w2 1 \u2264 2w2 0 w2 1+2w2 0\u2212w2 1 = 1. (12) This implies \u03b3 \u2208 [0, \u03c0/4]. We use this bound with (10) to arrive at the same bound on squared distance as (8) describes. Thus, the bound (9) applies for this case as well. 4) Both robots move: Fig. 1(c) shows robot 0 moving to point x and robot 1 moving to point y. The distance between robots changes from lt = \u2016r0 \u2212 r1\u2016 to lt+1 = \u2016x \u2212 y\u2016. Both robots move equal distance \u2016x \u2212 r0\u2016 = \u2016y \u2212 r1\u2016 = \u03b5. We will use additional notation a = \u2016O \u2212 r0\u2016 and b = \u2016O \u2212 r1\u2016. Using law of cosines l2t+1 = (a \u2212 \u03b5)2 + (b \u2212 \u03b5)2 \u2212 2(a \u2212 \u03b5)(b \u2212 \u03b5) cos \u03b4 = l2t + 2\u03b5(1 \u2212 cos \u03b4)(\u03b5 \u2212 a \u2212 b). (13) We start by bounding \u03b4. Note that \u03b3 + \u03b2/2 + \u03b4 = \u03c0. From (4) follows that 0 moves only when w1 + 2w0 cos \u03b1 \u2265 0. Given positive weights, if cos \u03b1 \u2265 0, then w1 + w0 cos \u03b1 > 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000219_iecon.2008.4758458-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000219_iecon.2008.4758458-Figure4-1.png", + "caption": "Fig. 4. 5-phase PM motor.", + "texts": [ + " (1) Arbitrary current waveforms satisfying equation (1) is shown in Fig 3(a). It can be seen form this figure (and from equation 1) that the currents in phase-m and phase-k have mirror symmetry with respect to the fault axis (in this case: phase-1). For only sinusoidal excitation currents, equation (1) can be expressed as: )cos( \u03b8\u03c9 \u2212= tIik , )cos()cos( \u03b8\u03c9\u03b8\u03c9 +=\u2212\u2212= tItIim . (2) The sinusoidal currents with mirror symmetry with respect to the axis of fault (phase-1 axis) are shown in Fig 3(b). A two pole, 5-phase PM machine is depicted in Fig 4. Considering only third harmonic components, the trapezoidal back emf (Fig 1) induced in the stator phases ( edcba eeeee ,,,, ) can be defined as in equation (3): )5/43cos(.)5/2cos(. )5/23cos(.)5/4cos(. )5/23cos(.)5/4cos(. )5/43cos(.)5/2cos(. )3cos(.)cos(. 31 31 31 31 31 \u03c0\u03c9\u03c0\u03c9 \u03c0\u03c9\u03c0\u03c9 \u03c0\u03c9\u03c0\u03c9 \u03c0\u03c9\u03c0\u03c9 \u03c9\u03c9 \u2212++= +++= \u2212+\u2212= ++\u2212= += tEtEe tEtEe tEtEe tEtEe tEtEe e d c b a (3) where \u03c9 is the rotor angular velocity, E1 and E3 are the amplitudes of the fundamental and third harmonic back-emf components. Assume an open circuit fault has occurred in the phase-a. Under this fault condition, the remaining four phases should be controlled to compensate the loss of the phases and produce the desired output torque. From fig 4, it can be noticed that with respect to the fault axis (phase-a axis), phase-b and phase-e are symmetrically located in space. Similarly, phase-c is positioned symmetrically with respect to phase-d. So, by applying equation (2), the unknown fundamental and third harmonic currents in the healthy phases, for fault tolerant operation, can be defined as in equation (4). )3cos(.)cos(. )3cos(.)cos(. )3cos(.)cos(. )3cos(.)cos(. 0 31311111 32321212 32321212 31311111 \u03b8\u03c9\u03b8\u03c9 \u03b8\u03c9\u03b8\u03c9 \u03b8\u03c9\u03b8\u03c9 \u03b8\u03c9\u03b8\u03c9 +++= +++= \u2212+\u2212= \u2212+\u2212= = tItIi tItIi tItIi tItIi i e d c b a (4) It can seen that, in this case, there are total eight unknown variables (I11, I12, I31, I32 \u03b811, \u03b812, \u03b831, \u03b832) which are to be determined for solutions of the fault tolerant currents" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure11-1.png", + "caption": "Fig. 11. Vehicle riding over bumps.", + "texts": [ + " Two models of the vehicle are considered in the study that follows: one with ideal kinematic joints in the suspen- sion systems and another using bushing joint in selected suspension elements. The location of the bushing joints on the rear suspension system, referred to as RB3, is represented in Fig. 9. The locations of the front suspension bushing joints are shown in Fig. 10(b) and c. The stiffness curves for the bushing joints model are obtained by a finite element model analysis, being the revolution bushing joints RB1, RB2 and RB3 the normal, tangential and rotational stiffness functions presented in Fig. 5. The application scenario, represented in Fig. 11, considers the vehicle riding over ten bumps, with a height of 0.1m. This case excites the roll motion of the vehicle chassis, making possible an evaluation of the suspension efficiency to reduce this chassis motion. Forward vehicle speeds of 60, 90 and 120Km/h are used to study the case. The dynamic behavior of the vehicle models for this scenario can be characterized by its vertical position and by the roll accelerations, depicted in Figs. 12 and 13. Analyzing the results presented in Fig. 12 for the vertical displacement of the center of mass of the chassis, it is observed that the vertical displacement is larger for the forward vehicle velocity of 120 km/h and smaller for the velocity of 60 km/h, as expected", + " The stiffness functions, which represent the perturbed functions, are increased by 1% for the sensitivity calculations. It must be noted that the damping coefficients remain the same in the reference and perturbed cases, i.e., equal to 0.01. The sensitivity analysis is performed to evaluate the influence of the bushing stiffness on the vehicle ride. The radial, tangential and angular stiffness represent independent design variables. The scenario in which the sensitivity analysis is carried consists in a vehicle traveling over several bumps with a forward speed of 90 km/h, as shown in Fig. 11. The reference functions for this scenario are the vertical position, vertical acceleration and roll acceleration of the vehicle chassis center of mass. The sensitivity of the vertical position of the vehicle mass center is presented in Fig. 15. It is verified that the vertical position is very sensitive to changes of the normal front bushing stiffness. In this case it is verified that it is necessary the increase of the damping coefficient in order to obtain a smooth the dynamic response. Fig. 15 also shows that the vertical chassis position is more sensitive to the tangential front bushing stiffness, since that the sensitivity values are positive and higher in the entire simulation time domain" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.38-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.38-1.png", + "caption": "Fig. 2.38 The independent double wheel escapement", + "texts": [ + " for production since 1980s till the present day. A sample is shown in Fig. 2.37. The newest escapement is the dual Ulysse escapement invented by Dr. Ludwig Oechslin in 2004 (Ludwig Oechslin 2004; Timebooth 2011). Dr. Oechslin received his Ph.D. in 1983 and his master watchmaker title in the subsequent year. Presently, he is the curator of the Mus\u00e9e International d\u2019Horlogerie, in La Chauxde-Fonds, Switzerland. The dual Ulysse escapement is perhaps inspired by the independent double wheel escapement invented in 1800s. Figure 2.38 shows the model of the double wheel escapement. Like many old designs, the independent double wheel escapement was abandoned because of its complexity and lack of reliability. As shown in Fig. 2.39, the dual Ulysse escapement consists of a balance wheel with a plate and a hairspring, a triangle-shape lever with two horns and two recesses and two escape wheels. There are also two pins used to limit the swing of the lever. Its most notable feature is the two escape wheels with specially designed tooth profile" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003583_iros.2011.6094491-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003583_iros.2011.6094491-Figure4-1.png", + "caption": "Fig. 4. A 2D plane on which the rimless wheel model is defined. The plane is determined by two vectors: the vertical through the CoM and the velocity T G of the CoM. pI is the actual CoP of the robot and P is the projected CoP located on the wheel plane about which the wheel rotates. The region enclosed by the dotted black lines is the support convex hull for the robot. The two feet of the robot are exaggerated to clearly show the convex hull and the points.", + "texts": [ + " Note that we first compute the GFPE and the corresponding configuration of the rimless wheel will be created with its leg angle a defined as the half of the angle between the two spokes (See Fig. 6), which is computed according to the GFPE. Since the rimless wheel model is defined in 2D and the robot exists in 3D, we determine a plane on which the simplified rimless model resides. We assume that the robot will move in the same direction as the velocity of the robot CoM, Te, just after a push. Consequently, the GFPE will be located on the line of Te projection on the ground. As shown in Fig. 4, the 2D plane for the rimless wheel is defined by two vectors: Te and the vertical through the CoM. Note that the magnitude and direction of the push are not required. The selected plane is solely dependent on the states of the robot after the push. Upon the 2D plane, we define an anchor point where the rimless wheel touches the ground and rotates about. Setting the anchor point is important since the robot has extended feet while the rimless wheel model has a point contact. Furthermore, the location of the anchor point determines the GFPE location and even the decision for taking a step. The previous works which used the FPE [11], [13] did not explicitly explain this choice of this point. They seemed to use the ground projection of the CoM, or the feet of their robot actually had a point contact. However, if we set the projected CoM as the anchor point, even a small push may result in a big step which may not be necessary at all. For instance, let a humanoid pushed from behind as in Fig. 4. If the anchor point (P in Fig. 4) is ahead of the projected CoM, the kinetic energy from a small push would be dissipated before the CoM rotates and passes over the vertical line through the anchor point, which implies no step is necessary since the robot would not topple forward from that small push. In contrast, if we use the projected CoM as the anchor point, the same push will make the CoM of the rimless wheel pass over the vertical, which would imply that the robot would topple forward and stepping would become necessary. In this work, we use projected CoP P as the anchor point, as shown in Fig. 4. We project the CoP p' of the robot on the 2D plane and assume that the rimless wheel rotates about the projected CoP, P. From now on, P represents the projected Cop. Note that the CoP reacts to a push much faster than the CoM since it depends on CoM acceleration. For example, just after a push from behind, the CoP will be ahead of the CoM as in Fig 4. Assuming that the robot knows the ground slopes, this GFPE exists in most cases unless stepping is physically not feasible due to the fixed length of the spokes or a steep slope (discussed in Section III-C). Note that most of the ground reference points for stepping such as FPE, ZMP [14], FRI [15] and capture point [8] are defined only on the fiat ground. One can envision a number of initial scenarios of push recovery after a push from behind, and the most probable one is shown in Fig. 5 where the robot is standing still and is subjected to a push from behind" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002588_1.4005215-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002588_1.4005215-Figure1-1.png", + "caption": "Fig. 1 Definition of preload, set bore clearance, and pad configuration", + "texts": [ + " In their model, the bump foils are modeled as primary surface plate-pin heat exchanger, and temperature distribution of the bump foils and cooling air along the axial direction were predicted using measured thermal contact resistance between the bumps and their mating surfaces. Lee and Kim [37] also present a 3-D THD model for double-acting foil thrust bearings considering cooling effect of the thrust runner disc by the cooling air plenum. Most THD studies on radial foil bearings are for single pad circular AFB while many industrial foil bearings have multipad configuration. Especially, AFBs with three pads with hydrodynamic preload [9,13] as described in Fig. 1 have many attractive features such as higher rotor-bearing stability and lower start friction compared to the single pad bearings. The preload Rp in Fig. 3 is defined as the physical distance between the bearing center and the top foil pad center when the top foil is assumed to follow the ideal circular profile as attached to the bump foil. The set bore clearance CS defines the minimum clearance between the rotor and bearing at the center of each top foil. When multiple top foil pads are used, the thermal boundary condition at the leading edge of each top foil is determined from mixing behavior of cooling air with exited air from the upstream trailing edge" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003472_j.cirpj.2011.01.010-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003472_j.cirpj.2011.01.010-Figure3-1.png", + "caption": "Fig. 3. Effect of substrate roughness on the thermal adhesion mechanisms of thermal spray coatings [17,19].", + "texts": [ + " These voids reduce the actual coating\u2013substrate contact area and thus bond strength. Surface asperities with an average distance that is approximately equal to the diameter of spray particles \u2013 with a typical size distribution from 10 to 60 mm \u2013 reduce the tendency of splats to spread out once they hit the substrate. Sobolev [19] suggested that splats with aspects ratios close to unity may increase the duration of the thermal pulse upon impact onto the substrate and, hence, the opportunity to develop diffusion bonding (Fig. 3). Grit blasting as a substrate preparation method offers a limited range of surface geometries in comparison to deterministic material removal processes. Another drawback of grit blasting is that grit particles may remain embedded on the substrate. Grit contamination has been reported to initiate delimitation cracking of thermal spray coatings [30]. With the exception of high pressure water jets [31], no alternatives to grit blasting have been evaluated in the scientific literature regarding the development of diffusion bonding of thermal spray coatings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001154_cec.2009.4983199-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001154_cec.2009.4983199-Figure6-1.png", + "caption": "Fig. 6. Representation of a sharp angle", + "texts": [ + " In other words, ( )L k normally calculates the length of chromosome and in case that collides with obstacles, it will add a large positive value so the chromosome is discarded. 0 if it collide with obstacles b 1 otherwise = (3) To find ( )S k , we calculate the angle of all genes except first and last ones and count how many of them are sharp turns. The angle of ip , i\u03b8 , can be calculated from the inner product of 1i ip p\u2212 and 1i ip p + which then has to be compared with a predefined angle 0\u03b8 as shown in Figure 6. For ip , a binary value is defined by (4) to mark it as a sharp turn. 0 0 0 1 i i i if if \u03b8 \u03b8 \u03b1 \u03b8 \u03b8 \u2264 = > (4) Thus, the number of sharp turns is the total number of sharp angles in a chromosome which is defined by (5). 2 1 ( ) n i i S k \u03b1 \u2212 = = (5) The evolutionary process of a genetic algorithm consists of evolutionary operators and an evaluation function which will result in a collision free, optimal or near optimal path. First we generate n CBPRM(c) paths as the initial population. The parameter n is a function of environment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001922_0954410011417510-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001922_0954410011417510-Figure1-1.png", + "caption": "Fig. 1 Missile axes and dynamic variables", + "texts": [ + " Simulation results are presented in order to illustrate the behaviour of the loop in terms of specification fulfilments and robustness to uncertainty. 2 PROBLEM DESCRIPTION Consider the tail-controlled pitch axis missile airframe. The control objective of the missile autopilot is to make the required acceleration track-given guidance command and must meet required performance criteria characterized by settling time, rising time, accuracy, and perturbation rejection. 2.1 Non-linear missile modelling A non-linear model for the missile motion in the pitch plane is taken from the previous work [17] and the variables are defined in Fig. 1. Equations of motion in the pitch plane are given by _ \u00bc QS mV Cz , Mm\u00f0 \u00de \u00fe Bz \u00f0 \u00de \u00fe q _q \u00bc QSd Iyy Cm , Mm\u00f0 \u00de \u00fe Bm \u00f0 \u00de \u00f01\u00de where , q, Mm, and are angle of attack, pitch rate, Mach number, and control fin deflection angle, respectively, and m, V , Iyy , Q, S, and d are mass, velocity, pitching moment of inertia, dynamic pressure, reference area, and reference length, respectively. The aerodynamic coefficients in equation (1) are represented as the function of Mach number and angle of attack. Bz \u00bc b1Mm \u00fe b2 Bm \u00bc b3Mm \u00fe b4 Cz , Mm\u00f0 \u00de \u00bc z1 \u00f0 \u00de \u00fe z2 \u00f0 \u00deMm Cm , Mm\u00f0 \u00de \u00bc m1 \u00f0 \u00de \u00fe m2 \u00f0 \u00deMm z1 \u00f0 \u00de \u00bc h1 3 \u00fe h2 j j \u00fe h3 z2 \u00f0 \u00de \u00bc h4 j j \u00fe h5 m1 \u00f0 \u00de \u00bc h6 3 \u00fe h7 j j \u00fe h8 m2 \u00f0 \u00de \u00bc h9 j j \u00fe h10 \u00f02\u00de where bi and hi are the constant coefficients" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002805_0278364912458463-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002805_0278364912458463-Figure3-1.png", + "caption": "Fig. 3. The basis function for pairwise leg repulsion, \u03c6, defined on the 2-torus. A minima appears along the cycle where legs are maximally separated in phase.", + "texts": [ + " In Klavins and Koditschek (2002), pairwise repulsion is formed using the simple trigonometric function, \u03c6( \u03c1i, \u03c1j) = cos[2\u03c0 ( \u03c1i \u2212\u03c1j) ]. This function achieves a smooth maximum value when legs are in phase, \u03c1i \u2212 \u03c1j = 0, and a smooth minimum for out-ofphase pairs, \u03c1i \u2212 \u03c1j = 0.5. By contrast, we have replaced this with an analogous function that produces smooth, differentiable minima, but has \u2018sharper\u2019 (continuous but not smooth) maxima at the cycle \u03c1i \u2212 \u03c1j = 0: \u03c6( \u03c1i, \u03c1j) = 1 \u2212 sin[\u03c0 ( \u03c1i \u2212 \u03c1j mod 1.0) ]. (9) The function \u03c6 is depicted as a surface over T2 in Figure 3. The gradient field associated with the resulting potential (8) is well defined over the complement of the pairwise in-phase subspace (a discontinuity in its gradient occurs along \u03c1i = \u03c1j, motivated by our desire for rapid transient dynamics). We apply this two-dimensional function to the entire torus by recourse to a hybrid tie-breaking rule that chooses a set of component functions, via c, in order to follow \u2018sensible\u2019 directions of convergence. While \u03c6 suffers from discontinuities in its gradient along the diagonal subspace of T2, we are amply repaid with a potential function that is convex over the complement of the diagonal subspace, a property that will be shown to be critical to our approach, as we now discuss" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000726_icems.2009.5382754-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000726_icems.2009.5382754-Figure3-1.png", + "caption": "Fig. 3. Basic model of segment type SRM.", + "texts": [ + " Figure 2 shows the torque-angle characteristics of the segment type SRM comparing with the same sized VR type SRM. It is shown that the maximum torque of the segment type SRM is twice as that of the VR type SRM. The average torque of the segment type SRM is 5.73N and the VR type SRM is 4.10N. However, the torque ripple factor of the segment type SRM is large, 199 %, compared with the VR type SRM, 146 %. Here, the torque ripple factor is calculated by the equation (1). %100 torqueAverage torqueMin.- torqueMax. factor ripple Torque \u00d7= (1) III. PHASE-NUMBER AND EXCITING CURRENT Figure 3 shows a basic model of the segment type SRM. \u03c6 is half of the slot pitch of the stator and \u03b4 is pitch between two adjacent rotor segment cores. Figure 4 shows excitation current waveform of the 3-phase segment type SRM. Each phase coil is excited in turn. Since the torque is small at start and end of excitation, the torque ripple becomes large. It can be seen on the torqueangle characteristics of the Figure 2. The excitation current waveform of the segment type SRM is uniquely decided depending on a combination of the number of phases and the number of segment cores. If we change the number of phases and the number of cores and additionally excite two or more phase windings at start and end of excitation, the torque waveform will be improved. In the Figure 3, the \u03c6 is given by the next equation. qm2 2\u03c0\u03c6 = (2) Where m : number of phases q : number of slots per phase The phase windings are excited for the interval \u03c6 as shown in Figure 4. In the Figure 3, the \u03b4 is given as follows. N \u03c0\u03b4 2 = (3) Where N : Number of segment cores The rotor rotates the distance of \u03b4 at one cycle of an electrical angle of the excitation current shown in Figure 4. In Figure 4, the \u03b1 is a distance for excitation from a certain phase to the next phase, and it is given as follows. mN \u03c0\u03b1 2 = (4) The relations between the number of phases m and the number of segment cores N are given by the next expression. mqqN \u2264+ (5) These equations show that \u03b1 can be shorten rather than \u03c6 by increasing the N greater than 3 and the m greater than 4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002614_j.phpro.2012.02.004-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002614_j.phpro.2012.02.004-Figure4-1.png", + "caption": "Figure 4. contact stresses in the inner ring and roller", + "texts": [], + "surrounding_texts": [ + "contact stresses are calculated. The stress distributions calculated are shown by Figs. 5 and 6. Zhang Yongqi et al. / Physics Procedia 24 (2012) 19 \u2013 24 23 Author name / Physics Procedia 00 (2011) 000\u2013000 Figure 7 shows that after the force of inertia taking into account, the contact stress 10% higher than the case without inertia force, so the main gear axle inertial forces at work can not be ignored. 24 Zhang Yongqi et al. / Physics Procedia 24 (2012) 19 \u2013 24 Author name / Physics Procedia 00 (2011) 000\u2013000 Ackonwledge This paper is supported by Graduate Innovation Fund of Jilin University (No.20101024)." + ] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure7-1.png", + "caption": "Fig. 7. Front suspension of the small family car.", + "texts": [ + " The multibody model of a small family car, presented in Fig. 6, is used to present the developments reported in this paper and to serve as the object to carry the sensitivity analysis on the effect of the bushing properties on the vehicle dynamics. The data required to build the multibody model is obtained through direct measurement of the real vehicle components and, therefore, the manufacturer has no responsibility on the data presented or used here. The vehicle front suspension is a McPherson mechanism, presented Fig. 7. In the rear suspension, presented in Fig. 8, a torsion beam suspension system is used, which is a common choice for this vehicle segment as it insures the compactness required for a small car while reducing the need for an anti-roll bar. However, the vehicle considered here still includes the anti-roll bar. The inertia properties of the rigid body, center of mass location and body fixed frame orientations, and all data for the suspensions, tires and stabilization bars, are described in the work by Verissimo [9]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002273_1350650111403580-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002273_1350650111403580-Figure1-1.png", + "caption": "Fig. 1 Schematic geometries of circular and lobed bearings", + "texts": [ + " With regards to micropolar lubricated multi-lobe bearings, as far as the authors are aware, there are only a few works [37\u201339] that also opt for a fixed preload value. Hence, the object of this study is to investigate the effect of the preload factor on the static performance of multi-lobe bearings lubricated with micropolar fluids. This will complement other works on micropolar lubricated bearings which has been reported in literature. The bearing configurations undertaken for this study are two-, three- and four-lobe finite bearings of symmetric geometries shown in Fig. 1. A non-circular lobed bearing is essentially an assembly of two or more partial circular arc bearings referred to as the lobes. The performance characteristics of the each partial bearing are then assembled, to obtain the overall bearing characteristics. A schematic diagram of a self-acting three-lobe journal bearing configuration and geometric space between journal and ith lobe, along with the coordinate systems used in the analysis, are shown in Fig. 2. Proc. IMechE Vol. 225 Part J: J. Engineering Tribology at UNIV OF PITTSBURGH on June 21, 2015pij" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003172_01691864.2013.776940-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003172_01691864.2013.776940-Figure3-1.png", + "caption": "Figure 3. Positional and angular errors.", + "texts": [ + " Finally, we connected two of these controllable wheels on the left and right sides of the main frame with rotary encoders (ME-30-P, Microtech Laboratory). A mobile PC (FMV-BIBLO LOOX U/G90 N, Fujitsu) was used for image processing, gait data acquisition, and user operation. A web camera (MCM-15 W, Loas) was connected to the PC via a USB port. We used a white tape to create a target line on the floor, which served as a guideline for walking. The camera captures the anteroinferior view of the floor and processes the image to detect the guideline.[13] The positional error (\u0394x in Figure 3) and angular error (D/ in Figure 3) are computed and sent to the main control box via a USB/serial converter. The sampling frequency used for image processing was 100Hz. A gait sensor (laser range finder, URG-04LN, Hokuyo Automatic) is also connected to the mobile PC via a USB port. The gait sensor was used to measure the stride width and step length of a subject (Figure 4). Within the measurement area (trapezoidal area in Figure 5), the gait sensor measures two independent Figure 2. Signal flow of i-Walker. D ow nl oa de d by [ K or ea U ni ve rs ity ] at 0 3: 01 0 3 Ja nu ar y 20 15 objects that are defined as the legs of the user", + " Velocity v [m/s] and _/ [rad/s] can be calculated as follows: v \u00bc 1 2 (rxl \u00fe rxr) (2) _/ \u00bc 1 L (rxl rxr) (3) If the current value of the velocity is less than or equal to an upper limit (vlim) [m/s], it is defined as the reference input of v (vref) [m/s]. If not, the reference input is set equal to vlim. The upper limit vlim depends on the user and is set to prevent users from falling. On the other hand, the reference input of _/ ( _/ref ) [rad/s] is set to reduce both positional and rotational errors (see Figure 3). Therefore, vref \u00bc v; if v 6 vlim vlim; else (4) _/ref \u00bc kxDx k/D/ (5) where kx= 2.0 [rad/(s m)] and k= 2.0 [l/s] are proportional gains for the position and orientation errors, respectively. Thus, the reference inputs for the rotational velocities of the wheels can be calculated as follows: xlref \u00bc 2vref \u00fe L _/ref 2r (6) xrref \u00bc 2vref L _/ref 2r (7) where \u03c9rref and \u03c9lref denote the references of the rotational velocities of the right and left wheels, respectively. Finally, the braking torque T [Nm] of each wheel can be defined as follows: Tn \u00bc kx(xn xnref ); if xnref 6 xn 0; else (8) where the subscript n takes the value l for the left wheel and r for the right wheel" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003551_iros.2011.6094417-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003551_iros.2011.6094417-Figure5-1.png", + "caption": "Fig. 5: Simplified kinematics of the head", + "texts": [ + " (1) Excluding the boundary values \u00b1\u03c0/2 of the tilt angle, the adopted description does not suffer any singularity. The pointing angles are evaluated as: \u03b8pan = atan ( gy gx ) , \u03b8tilt = atan \u239b \u239d gz\u221a g2x + g2y \u239e \u23a0 . (2) The pointing direction n of the robot head is chosen as the x axis of the end-effector (camera) frame, and is determined once the kinematic model of the head is established. In order to avoid the implementation of an on-line calibration of the stereo system, motions of the two eyes are kept parallel. Thus, a simplified kinematic scheme of the head with only 6 dof has been considered (see Fig. 5), without loss of the generality for the proposed control scheme. The direct kinematics of the head is defined by the homogeneous matrix BTE(q) = BT 0 0T 6(q) 6TE = ( n(q) s(q) a(q) p(q) 0 0 0 1 ) , (3) where 0T 6(q) is calculated using the Denavit-Hartenberg parameters in Tab. I, BT 0 is a (constant) rotation by \u2212\u03c0/2 around the xB axis, and 6TE = I . The head gazes perfectly at the target when n(q) = g. Evaluating n(q) from eq. (3) and setting g = n(q) in eq. (2), in the spirit of the task function approach (see, e" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002028_j.ymssp.2012.05.013-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002028_j.ymssp.2012.05.013-Figure2-1.png", + "caption": "Fig. 2. Test gearbox.", + "texts": [ + " Generally, accessories for gas turbine engines can be divided into two categories: (1) those driven by bleed air taken from the compressor section of the engine and, (2) those driven mechanically by an accessory gearbox connected directly or indirectly to the engine main shaft. In general, the mechanical connection from the engine shaft may be through an engine-mounted gearbox or through a power take-off shaft to a remotely mounted gearbox. Fig. 1 presents a cutaway section of a J85 engine showing its main components as well as the accessory gearbox attached to the engine casing below the compressor section. The gearbox is characterized by four parallel shaftgears (Fig. 2) driving the various accessories mounted at each pad, as shown in the detailed schematic of Fig. 3. This gearbox variant is driven axially (axis B) via a transfer gearbox connected to the engine by an output shaft (axis A). Power is then transmitted directly to the over-speed governor and indirectly to the remaining accessories (starter/generator, hydraulic, fuel and lubrication pumps) through gears F, C and D. Rolling-element bearings are used to support and guide all shafts as shown in Fig. 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002238_elan.201100540-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002238_elan.201100540-Figure7-1.png", + "caption": "Fig. 7. Image of the prototype (a) and the schematic diagram of the electronic circuit of the prototype (b).", + "texts": [ + " As storage conditions, the biosensor was maintained dry at room temperature. As shown in the Figure 6a, the enzymatic activity decreases after about 1 month; however, up to the period tested, the biosensor retained the original sensitivity for paraoxon detection (Figure 6 b). After the development and the analytical characterization of BChE biosensor based on screen-printed electrode made only with the reference and working electrodes (Figure 2), the biosensor was integrated in the prototype (Figure 7 a). The prototype was developed in conjunction with Strictes+AeG Company and composed of a cell described in detail in Experimental (Section 2.2 and showed in Figure 1). The biosensor was inserted in the cell and the substrate solution was added by using a syringe (Figure 1 b). The system was able to detect the nerve agent in gas phase because the device is composed of a little ventilator/fan able to sampling 20\u201325 L of air/min, and a canal that connects the fan with the electrochemical cell, in order to gurgle the air sampled by the fun into the electrochemical cell, thus in the solution in which the biosensor is in contact. The prototype is composed also of an electronic circuit able to control the fun, apply the potential, register the signal and give the alarm. The electronic circuit, as schematized in the Figure 7 b, is supplied by a battery and a voltage regulator in order to stabilize the potential at 3.3 V. The circuit is composed of: \u2013 an analog section, that includes 2 stages: i) Applied potential (+200 mV) to the sensor+ input I/ V amplifier which converts the input current signal in output voltage signal; ii) Low-pass filter to cut-off high frequency interference; \u2013 a digital section that includes 1 stage: i) A/D converter, input stage of the microcontroller, thus the final output signal can be the on or off of the alarm and in the first case, the alarm led becomes red" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002124_j.triboint.2010.05.005-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002124_j.triboint.2010.05.005-Figure1-1.png", + "caption": "Fig. 1. Schematic illustrat", + "texts": [ + " Recently, a ball-on-disc optical test rig in the author\u2019s lab has been updated to incorporate spinning into EHL contacts by the contact center offset from the disc rotation axis and measurements under pure disc sliding conditions have been reported [17]. In this paper, pure rolling configuration of ball-ondisc contact are used and the influence of spinning on film thickness and the film shape were investigated. Numerical calculations have also been completed to justify the experiment results and gain more insights into the film formation. Fig. 1 shows a schematic illustration of the optical EHL test rig used. The synchronous pulley A is driven by a serve-motor. The glass plate C rotates around its axis O1O2 at an angular speed o. The steel ball D is loaded against the glass plate C, and can rotate freely around its axis O3O4. The offset r of the Hertzian contact center with respect to axis O1O2 can be adjusted precisely by the XY table E. Fig. 2 presents the ball-on-disc configuration in more details. The velocity profile in the contact region along the y-axis is given in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001203_j.na.2009.01.238-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001203_j.na.2009.01.238-Figure1-1.png", + "caption": "Fig. 1. Differential driving mode for an 8-pole AMB.", + "texts": [ + "238 Assuming there is no eddy current losses, magnetic flux leakage, hysteresis, saturation and also relative permeability of the material \u00b5r 1, electromagnetic force is given as [12], f = 1 4 \u00b50N2A i2 s2 cos\u03b1 (1) where \u00b50 is the permeability of vacuum, A is the cross sectional area of the electromagnet, N is the number of windings, s is the nominal air gap between the rotor and the stator and \u03b1 is the angle between stator pole and its effecting axis. This equation is derived only for 2-pole magnetic bearing element but the AMB used in this paper is an 8-pole magnetic bearing system which works in differential driving mode. This working principle is depicted in Fig. 1. Here the current ib is called bias current. Bias current is supplied constantly to each 2-pole element. In the differential driving mode, control current is added to the bias current while in the opposite direction it is subtracted from it. In the positive directions the gap between the rotor and the AMB is assumed to be decreasing while the resultant current is the summation of bias current and regulating current. Ignoring the geometric coupling between horizontal and vertical axes net forces acting on the rotor in positive and negative directions are Fx+ = 1 4 \u00b50N2A cos\u03b1 (ib + ix)2 (s\u2212 x)2 (2) Fx\u2212 = 1 4 \u00b50N2A cos\u03b1 (ib \u2212 ix)2 (s+ x)2 ", + " Mx\u0308G = Fx1 + Fx2 + Fxu +Mg (5) Ix\u03b8\u0308y = \u03c9Iz \u03b8\u0307x + Fx1L1 \u2212 Fx2L2 (6) My\u0308G = Fy1 + Fy2 + Fyu (7) Iy\u03b8\u0308x = \u2212\u03c9Iz \u03b8\u0307y \u2212 Fy1L1 + Fy2L2, (8) where Ix and Iy are transverse mass moment of inertia about x- and y-axes respectively, Iz is polar moment of inertia about z-axis,M denotes mass of rotor, L1 and L2 are the distances between the centre of mass and magnetic bearings on the lefthand side and on the right-hand side respectively. Fxu and Fyu are the components of force created due to mass unbalance in horizontal and vertical axis respectively. Mass unbalance is shown in Fig. 1. The following equations are the components of unbalance force: Fux = M\u03c92r cos(\u03c9t) (9) Fuy = M\u03c92r sin(\u03c9t). (10) The dynamical equations of the rigid rotor are derived with respect to the geometric centre. Since the rotor is supported by its two ends and it is assumed that the sensors are located in the middle of magnetic bearings, it is necessary to rewrite the equations in terms of bearing coordinates. Using trigonometric relations and assuming small oscillations of \u03b8 (i.e. sin(\u03b8) \u223c= \u03b8 ), the following equations are obtained: x1 = xG + L1\u03b8y y1 = yG + L1\u03b8x (11a) x2 = xG \u2212 L2\u03b8y y2 = yG \u2212 L2\u03b8x" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003316_s12555-012-0335-3-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003316_s12555-012-0335-3-Figure1-1.png", + "caption": "Fig. 1. Helicopter and it\u2019s frame.", + "texts": [ + " In this paper, two distinct models of the yaw dynamics of the UAV helicopter are considered: a higher-fidelity simulation model (SM) is used exclusively for closed-loop simulation, whereas a reduced complexity control-design model (CDM) is used for control design and stability analysis of the closedloop system. Apart from differences in the actual models, the SM and CDM are similar to the so-called truth model and curve-fitted model (CFM) in [10]. 2.1. Simulation model The SM adopted in this study is the model developed by [22] for the yaw dynamics of a UAV helicopter. In this subsection, a framework of the simulation model for the helicopter (see Fig. 1) is set up using rigid body equations of motion of the helicopter fuselage. In this way the influence of the aerodynamic forces and moments working on the helicopter are expressed. The total aerodynamic forces and moments acting on a helicopter can be computed by summing up the contributions of all parts on the helicopter (including main rotor, fuselage, tail rotor, vertical fin and horizontal stabilizer). So, the yaw channel dynamic equations are given by: , , zz mr tr fus hs vf r I r N N N N N \u03d5 = = + + + + (1) where \u03d5 and r are the yaw angle and angular rate of the helicopter; zz I is the inertia around z-axis; , mr N , tr N , hs N fusN and vfN present the torque of main rotor, tail rotor, horizontal, fuselage and vertical fin worked on the helicopter respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001200_00405160903178591-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001200_00405160903178591-Figure2-1.png", + "caption": "Figure 2. Details of sliding friction.", + "texts": [ + " At the contact points, the molecules on the opposite sides of the surface are so close to each other that they exert strong intermolecular forces among the molecules. When one body is pulled across another, the frictional resistance is associated with the rupturing of these thousands of tiny welds that continually reform as new contacts D ow nl oa de d by [ Il lin oi s In st itu te O f T ec hn ol og y] a t 1 9: 39 0 2 M ay 2 01 3 Figure 1. Highly magnified view of a section of a finely polished steel surface. are made, as seen in Figure 2. It would be appropriate to divide the studies related to the friction into the following two groups: Friction of hard solids. Friction of fibrous and viscoelastic materials. Textile fibres mostly fall into the second category in which the friction of fibrous and viscoelastic materials is valid. D ow nl oa de d by [ Il lin oi s In st itu te O f T ec hn ol og y] a t 1 9: 39 0 2 M ay 2 01 3 A basic definition of friction given in physics books is \u2018the resistance encountered when two bodies are brought to contact and allowed to slide against each other\u2019 [1,2,11,12,20,24\u2013 27]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003692_978-3-642-23681-5_13-Figure13.17-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003692_978-3-642-23681-5_13-Figure13.17-1.png", + "caption": "Fig. 13.17 Location of pressure measuring points (Hargreaves et al. [16])", + "texts": [ + " For all tests reported here, the bearing material was an elastomeric material made from thermosetting resins which are three-dimensional cross-linked condensation polymers. The non-metallic bearing was mounted in a steel sleeve with an interference fit and the sleeve-bearing assembly was fixed into the housing. Pressures were measured along the groove located in the loaded region of the bearing and around the circumference of the bearing half-way along its length for various operating conditions\u2014see Fig. 13.17. There were 16 tapping points circumferentially and 7 axially. Special note was made of pressures near to the groove edges. Flexible tubes were fitted to each tapping point and taken to a manifold where the pressure was measured by a pressure transducer, one at a time. An air release valve was located near to the manifold. The load was applied by a dead weight to the top of the bearing housing thereby ensuring that the minimum film thickness was located near to the top of the bearing. Only bearings with three equi-spaced axial grooves along the complete length of the bearing were tested" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000857_1.2805443-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000857_1.2805443-Figure3-1.png", + "caption": "Fig. 3 Left: tangential components o ferential area \u201edA=rdrd \u2026 and x comp", + "texts": [ + " We have assumed a Gaussian distribution of asperity height sum so that IN h, s = h 0 2 s s\u2212h s \u2212 h \u2212 r2 2 s 3/2 1 + r2 s 2 \u22121/2 e\u2212s2/2rdrds 13 In Eqs. 12 and 13 , the parameters have been normalized with respect to the standard deviation of asperity height sum so that s is z / , h is d / , and r, , and s are now the normalized values using as the normalization parameter. Tangential Force The tangential components due to various interactions cannot be algebraically added as they are projections of contact force onto the mean plane and depend on circumferential position of asperities on surface S2 Fig. 3 . In considering the tangential component of contact force, we seek the components of the tangential contact force along an axis of interest, for instance, the tangential force component along the x axis, depicted in Fig. 3. Of course subjecting the x component to a statistical treatment would yield zero since the x components of the tangential force corresponding to positive contact slope would equalize those due to negative contact slope. We are interested in formulating the cumulative effect of x component of tangential force in the right side as shown in Fig. 3 . Hereafter as we generate result for the x component of the tangential force due to positive contact slope, ty contact we will refer to this as the \u201ctangential force\u201d and denote the force Transactions of the ASME of Use: http://www.asme.org/about-asme/terms-of-use c t d a t l t t s b f A J Downloaded Fr omponent Fx. By this, it should not be inferred or assumed that here is a net tangential force acting on an asperity per above iscussion. The goal here is to account for the tangential force an sperity would experience on each side and therefore accumulaion or summation of such forces would establish the tangential oad on a surface from each side, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001400_00423110903126478-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001400_00423110903126478-Figure1-1.png", + "caption": "Figure 1. (a) Suspension of the standard UIC Y25 bogie. 1-bogie frame, 2-inner spring, 3-outer spring, 4-friction piston, 5-sliding surface, 6-Lenoir link, 7-axle box, 8-pivot, (b) a photograph of the suspension.", + "texts": [ + " A dithered system of technical importance is a railway freight wagon with friction dampers developing two-dimensional friction in the primary suspension, where the M-F dither exciting dampers is generated by the rolling contact of wheel and rail. The M-F dither, by changing the properties of friction damping in the suspension, influences the ride dynamics of the wagon. *Email: jpt@simr.pw.edu.pl ISSN 0042-3114 print/ISSN 1744-5159 online \u00a9 2010 Taylor & Francis DOI: 10.1080/00423110903126478 http://www.informaworld.com D ow nl oa de d by [ O tte rb ei n U ni ve rs ity ] at 0 3: 30 2 1 A pr il 20 13 An example of the suspension with friction dampers exposed to dither is the suspension of the standard UIC Y25 freight bogie shown in Figure 1. The load transmitted by an inclined link 6 presses the cap 5 into the piston 4, which transmits the load on the vertical surface of the axle box. The axle box cannot move in the longitudinal direction under the action of the axle box force of moderate value because it is clamped between the piston and the guiding vertical surface of the frame. Only when the longitudinal axle box force is of sufficient value and acts in the direction towards the link, can the axle box displace longitudinally in this direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001154_cec.2009.4983199-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001154_cec.2009.4983199-Figure5-1.png", + "caption": "Fig. 5. usage of (a) crossover, (b) change, (c) smooth, (d) shortcut operators", + "texts": [ + " Evolutionary operators Our method uses ordinary crossover and mutation which are standard operators in genetic algorithms. In addition, we use change, smooth and shortcut operators as a subset of mutation in order to manipulate an individual. Detail of each operator is described below: 1) Crossover: This operator combines two individual to create a new chromosome. By applying this operator, two reference genes ( , )m n are selected randomly and chromosomes are swapped from m th and n th genes. Figure 5 (a) shows an example of this operator. 2) Change: This operator changes the Cartesian point of some selected genes from a chromosome to create a new mutated one. When this operator is applied, a random point is searched on a circle centered at current gene with a predefined radius (Figure 5 (b)). The radius can be a function of obstacles sizes or it can be a constant moderate value. 3) Smooth: This operator try to smooth sharp turns in an individual. With a predefined angle as sharp turn (say 0\u03b8 ), this operator finds the intersection of a circle centered at sharp angle ( ip gene) and two lines passed through this gene ( 1i ip p\u2212 and 1i ip p + ), so a new smoothed path is created. The radius of the circle like change operator can be a function of obstacles sizes or it can be a constant moderate value. Figure 5 (c) shows an example of this operator. 4) Shortcut: This operator will eliminate redundant genes from an individual. Figure 5 (d) shows an example of this operator. For each gene from beginning of chromosome we reversly check which gene is visiable from current gene, so we will be sure that if gene i sees gene j , ( 0 i j n\u2264 < \u2264 ), then all of the genes between i and j will be eliminated. 2009 IEEE Congress on Evolutionary Computation (CEC 2009) 2093 Next we have to decide which of the children are deserved to go to next generation. The two important metrics in the path quality are length and smoothness. In other word, the final path has to be minimized in length and number of sharp turns" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002070_1.1658083-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002070_1.1658083-Figure4-1.png", + "caption": "FIG. 4. Pressure exerted on the equivalent rigid wall by a ran dom distribution of identical defects according to the spring model. Motion within the limits -80 /2585 +80 /2 is reversible while motion outside these limits shows frictional hysteresis.", + "texts": [ + "= -NkzN2= -NWo, (10) where the minus sign indicates that the pressure is in the minus Z direction and the subscript s indicates that the pressure is saturated (constant). It is assumed that the total number of attached defects is large so that we may deal with average values and not be concerned with fluctuations from the average. The wall is now stopped and allowed to move backward a distance less than ZQ. This situation is shown in Fig. 3. The pressure Puis now unsaturated and depends on z. It is given by (+'0/2 Pu=Nk L (z'-z)dz'= -NkzQz= -Nfoz. -zo/2 (11) The pressure exerted by the defects on the wall is then as shown by the full line in Fig. 4. Wall motion within the limits -zo/2':::;z':::;+zo/2 is reversible and has no hysteresis. If the wall is moved past zo/2, say to point f3z= 2M.H+P(z) , (12) where f3 is a viscous-damping constant, H is the applied field, M. is the saturation magnetization, and P(z) is given in Fig. 4. It may be noted that Fig. 4 is identical with Fig. 5 of Ref. 8. As a consequence, all of the results derived in that paper follow from the present model. In particular, it predicts that (1) There is a reversible spike of wall velocity at the leading and trailing edges when a rectangular pulse of field is applied to a wall. (2) The steady-state velocity in a constant field is proportional to H - He, where H. is the threshold field for irreversible wall motion, as shown in Fig. 2(b). (3) The small-signal hysteresis loop is Rayleigh-like if the distribution of defect strengths is essentially constant for small values of zoo The present model is more satisfactory than that of Ref", + "201 On: Sat, 11 Oct 2014 19:06:19 MODEL OF MAGNETIC HYSTERESIS 2835 When the extension of the spring reaches a value Zo characteristic of the defect, it breaks and the energy stored in it is lost. The uniform motion of the wall is accompanied by a continuous attachment and breakage of springs. The result is a frictional energy loss. The picture shown in Fig. 1, which has frequently been used in discussing hysteresis, is seen to be not applicable to most materials. It must be replaced by JOURNAL OF APPLIED PHYSICS the result of generalizing Fig. 4 to cover a distribution of defect strengths as was done in Ref. 8. ACKNOWLEDGMENTS It is a pleasure to acknowledge the invaluable assistance of David Leedom and William Holsten who did the programming. VOLUME 40, NUMBER 7 JUNE 1969 Magnetic Relaxations and the Koops Inhomogeneity in Ferrites F. HABEREY* Institut fur Werkstoffe der Elektrotechnik der Rlteinisch-Westjiilischen Technisclten H ochschule Aachen, 51 Aachen, West Germany (Received 13 December 1968; in final form 3 February 1969) Conductivity measurements in various polycrystalIine ferrites have shown that the temperature de pendence of the conductivity has to be described by two different activation energies, one which is attributed to the bulk material and the other which is due to the poorly conducting layers introduced by Koops" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure10-1.png", + "caption": "Fig. 10. The scheme of the gear in extreme gearing position.", + "texts": [ + " The variations of the pressure angle\u03b1w are presented in Fig. 9, which shows that the number of teeth z1 has a significant influence, the pressure angle increasing emphaticallywith\u03c62 (or S1); frompractice, the angle\u03d52 should not exceed 40\u00b0 [7]. Introducing the limit \u03c62max=40\u00b0, the following optimal variants assure the required transmission ratio \u0393\u22451.3: z1/e=25/15;20/20;20/15. An image of the acceptable gear solution (z1=25,e=20mm,z2=24,m0=4mm), drawn in one of the extreme gearing position (end of stroke, S1=27mm,\u03c62=50\u00b0) is shown in Fig. 10. As a global observation, the numbers of teeth (z1, z2) must be quite close one to another (the translating wheel and the geared segment are almost equals), and the eccentricity (e) must be less than half of their radius. The graphs show a fairly large variation range of all functions, so that it generates no difficulty in finding the optimal solution when applying to the steering boxes of the vehicles. The optimal correlation of the parameters (z1, z2 and e) for a certain module (m0) is always possible for the variant of sector with negative eccentricity (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002124_j.triboint.2010.05.005-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002124_j.triboint.2010.05.005-Figure12-1.png", + "caption": "Fig. 12. Influence of speeds on the film thickness, w\u00bc16 N,Ssp\u00bc0.4, PB1300.", + "texts": [ + " Numerical calculations have been carried out to examine the relation of film thickness versus speed under spinning conditions. Theoretical speed indices under the conditions of Fig. 6 are obtained, 0.64 for hcen, 0.66 for hminO, 0.75 for hminL and 0.79 for hminR, and are different from those measured. However, both the experiment and the numerical work show that the speed index of the central film thickness are similar to that predicted by Hamrock\u2013Dowson formula, and those of the minimum film thickness at the side-lobes are larger. The lowest film thickness hminR has the largest speed index. Fig. 12 gives the calculated relations of hcen vs ue and hminR vs ue, and the measured data are also plotted together. Numerical computations present film thicknesses lower than that in the experiments. Fig. 13 gives the calculated side lobe film thicknesses hminL and hminR under different spin ratios, which show trends similar to those in the experiments of Fig. 7. However, both hminL and hminR in numerical calculations are different from those in the experiments, but change in similar way. As illustrated in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003460_1350650112466768-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003460_1350650112466768-Figure4-1.png", + "caption": "Figure 4. Constraint condition of piston.", + "texts": [ + "18 The compliance matrix can be generated using a symmetric finite element model (FEM) in the following manner: For nodes not on the symmetry line, two unit forces are applied symmetrically at xj and x0j, and the net displacement at each node, xi, is measured. For nodes along the symmetry line, a single unit force is applied at xs, and the displacement at each node, xi, is measured. These results are symmetric, and therefore data for only one half of the piston is recorded. The constraint condition of piston is as shown in Figure 4 and the generation of symmetric compliance matrix is as shown in Figure 5. This matrix is then used to calculate the radial deformation of the piston due to the normal forces generated due to the contact and film pressure. The contour of deformation due to a unit normal pressure located at one node is shown in Figure 6(a) Radial deformationf g \u00bc Compliance matrix\u00bd Normal forcef g \u00f021\u00de Pressure model. Combustion gas pressure can result in significant radial deformation of the piston, particularly close to combustion TDC" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000644_j.ymssp.2008.12.007-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000644_j.ymssp.2008.12.007-Figure2-1.png", + "caption": "Fig. 2. Geometry of the cracked shaft segment.", + "texts": [ + " from [1,7,8], since the stress intensity factors K Ii, K IIi, K IIIi are quadratic functions of the virtual generalized loadings Pi, the coefficients of flexibility obtained using (3) are functions of the material constants and cracked shaft cross-section parameters a and D. These coefficients create the 6 6 symmetrical local flexibility matrix. The inverse of this matrix becomes the crack local stiffness coefficient matrix of the following form: c11 0 c13 0 0 c16 0 c22 0 c24 c25 0 c13 0 c33 0 0 c36 0 c24 0 c44 c45 0 0 c25 0 c45 c55 0 c16 0 c36 0 0 c66 2 6666666664 3 7777777775 . (4) The above matrix is not diagonal. If we assume, according to Fig. 2, that the direction \u2018\u20181\u2019\u2019 corresponds to translational motion along 0Z axis, the direction \u2018\u20182\u2019\u2019 corresponds to rotational motion around 0z axis, the direction \u2018\u20183\u2019\u2019 corresponds to translational motion along 0z axis, the direction \u2018\u20184\u2019\u2019 corresponds to rotational motion around 0Z axis, the direction \u2018\u20185\u2019\u2019 corresponds to translational motion along the shaft rotation axis 0x and the direction \u2018\u20186\u2019\u2019 corresponds to rotational motion around 0x axis, it follows from (4) that due to local cross-sectional anisotropy caused by the crack the shaft transverse motion is directly coupled with its rotational motion around the rotation axis as well as the rotational motions around diameters are directly coupled with translational motions in the axial direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002510_1.3616719-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002510_1.3616719-Figure1-1.png", + "caption": "Fig. 1 Compressor rear f rame liner fatigue fa i lu re\u2014view looking forward (photograph no. 2 1 6 3 5 7 )", + "texts": [ + " With turbomachinery wheel space chambers, powerful currents of air may flow in and out of diffuser inlet. Efficient diff users Fig. 9 Compressor discharge annular d i f fuser\u2014where the compressor discharge seal is at a substantially smaller diameter than inner w a l l of main f low annulus, check for possible crossflow and cellular recirculating flows between w h e e l space chamber and diffuser inlet RADIAL STRUTS ANNULAR CHAMBER BETWEEN BLADE EXIT AND STRUTS,\u00a9\u2014\u2014 WHEEL SPACE CHAMBERS- INFLOW1 REGION n = 1 CIRCUMFERENTIAL WAVES n=3 CIRCUMFERENTIAL WAVES Fig. 1 0 Wheel space chamber coupled to diffuser inlet. Where an ax ia l gap connects whee l space chamber and diffuser inlet, cellular recirculating flows m a y occur. are known to be very sensitive to flows into and out of inlet, especially when this parasitic flow is injected perpendicular to or forward facing into the main stream. McMahan has thoroughly investigated the quantitative relationships [2]. The outer resonator of the Boys double resonator is open at one end. In turbomachinery, the \"organ-pipe\" resonator component is often closed at both ends" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001656_s11431-010-0064-x-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001656_s11431-010-0064-x-Figure6-1.png", + "caption": "Figure 6 Inputted model of computational example.", + "texts": [ + "i j i j i jn n t+ = + \u0394 \u22c5v v f If , ,( 1) ( ),i j i jn CFC+ \u2209v p the cutter orientation ,( 1)i jn +v will be pulled back to the feasibility cone CFC(pi,j) along the opposite direction of the force fi,j. Step 4. Calculate the force fi,j at each mesh point and the total force F(vn+1) in the spring mesh. The algorithm will terminate if 1( ) 0,n+ =F v 1( ) ( )n n+ >F v F v or ( )1 t( ) ( ) ( ) .n n n e+\u2212 = Otherwise, go to Step 2. The developed algorithm is implemented and validated by using the blade model shown in Figure 6. The blade can be machined by a ball-end cutter. Since there is a one-to-one relationship between the cutter center point and the CC point, a cutter center point mesh of the ball-end cutter is inputted. The cutter center point will not be changed to each cutter location data in the process of adjusting tool orientation and the CC point will not get shifted. For every cutter center point, the accessibility cones are first computed and the convex feasibility cones are approximated. The cutter orientations are then wholly optimized based on the cutter center point mesh", + " The tool length is 50 mm and the radius of its tool holder is 18 mm. There are 8 rows in the inputted cutter center point mesh. Each row consists of 211 cutter center points. The accessibility cones of all mesh points can be computed by using the GPU-based algorithm proposed in ref. [14]. 13.44 s are spent in checking the accessibility of 258 candidate cutter orientations. The feasibility cones can be obtained according to eq. (5). One of the feasibility cones is described by the triangle mesh in Figure 6. The basic step of the algorithm is to approximate the feasibility cone by a simple convex shape such as a conical cone. The method to obtain the conical cone is shown in Figure 7. A depth-first search of the triangle mesh is adopted to thin the feasibility cone. The center of the conical cone is determined by the deepest feasible orientations. For example, the center orientation vc of the conical cone is the feasible orientation with the depth of 3. The vertex angle of the conical cone is the minimal angle between the central orientation and the boundary of the feasibility cone. , ,( , ) ( , )i j i jCFC S FC S\u2286p p is guaranteed by this method. The obtained feasibility cone is shown in Figure 6 and can then be regarded as a continuous convex set. Based on the continuous convex feasibility cone, the proposed cutter orientation optimization model can be solved. Ref. [13] developed a method to select tool posture independent of previous posture and cutter orientations are central orientations of accessibility cones. By the similar method, initial cutter orientations are set as the central orientations of conical feasibility cones. In the initialization step, nmax=60, \u0394t=0.5 and et=10\u22124. The total force |F(v)| and the whole-smoothness measure m(p,v) with respect to iterative steps are shown in Figure 8" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000502_24748668.2008.11868439-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000502_24748668.2008.11868439-Figure3-1.png", + "caption": "Figure 3. Percentage use of types of jump in competition", + "texts": [ + " SportsCode 2007 (GameBraker) software was used for the study and the following stages were sequenced: A) Recording and Digitalisation of the images, B) Creation of a jump type matrix, C) Image capture for each matrix code, and D) Combination of the matrix codes to obtain jump performance in real beach volleyball competition. All the data was downloaded to a Microsoft Excel spreadsheet using the categories of jump type and time to calculate the frequency. The first set of results showed the mean use of the different types of jump by players in the four official competitive matches (Figure 2). The mean number of service jumps (SS) was 36.5\u00b13.2, the mean number of smash jumps (SSJ) 96.0\u00b13.4, and the mean number of block jumps (BJ) was 86.5\u00b12.3. The second set of results (Figure 3) of the comparison looked at the percentage use of the different game actions, with 44% for smash jumps (SSJ), 17% for service jumps (SJ) and 39% for block jumps (BJ). Later analysis related real playing time (TR) to the quantification of the jumps. In this section we extracted the real playing time per match, which is the difference between absolute playing time (TA) and the times that the ball was not in play, and the total number of jumps carried out (Figure 4). . The same was done with the mean real playing time per set and the corresponding number of jumps (Figure 5) and the mean real playing time per point and the number of jumps (Figure 6)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure8-1.png", + "caption": "Fig. 8 Negative volume of the grinding wheel and the volute casing (outer surfaces of the flow area)", + "texts": [ + " Furthermore, the SST-kx turbulence model is utilized, which offers good approximations for both areas inside the flow and boundary areas (e.g. pipe wall). The pump characteristic curve of the GIC (i.e. the total head versus the flow rate) is simulated by CFD. To evaluate the pumping action of the GIC, it is simulated with the same boundary conditions as common centrifugal pump impellers. For this reason, a volute casing is employed to deliver the fluid away from the coolant outlets. This volute casing delivers the fluid to an outlet branch (Fig. 8) and avoids boundary effects on the fluid flow inside the grinding wheel. The simulation results are the delivery head and the flow rate. In addition to the delivery head, also the torque can be simulated and hence the driving power of the grinding wheel can be calculated. The driving power corresponds to the spindle power and can be compared to experimental results. In addition to integral values, like delivery head and power, local values (pressure and fluid velocity at discrete points) can be simulated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001743_3.30282-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001743_3.30282-Figure2-1.png", + "caption": "Fig. 2 Orientation of solar sensor fields of view.", + "texts": [ + " This paper describes the onboard solar aspect sensing instrument, calibration and analysis techniques, data handling and reduction processes (manual and automated), accuracy limitations, results of recent firings, and limitations of and usefulness of this technique. D ow nl oa de d by U N IV E R SI T Y O F M IC H IG A N o n Fe br ua ry 1 9, 2 01 5 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /3 .3 02 82 The solar aspect angle a (Fig. 1) can be measured by a simple geometric arrangement of sensing cells and slits. Two silicon solar cells are mounted in fixtures which define almost planar fields of view (Fig. 2). Each field of view is established by a slit, serrated along its length to absorb internal light reflections, with reflective surfaces at each end of the slit to permit wide viewing in the plane. The angle 2ft between the two fixtures in a plane perpendicular to the projectile axis (Fig. 3b) is set to a predetermined value. As the projectile rolls, the sun is intercepted by each fixture in turn at a time when the solar vector lies in the plane of the field of view of that fixture. The time interval y between successive intercepts of the sun by the same cell fixture is the period of the rolling motion", + " The electronics on board the projectile and on the ground do not introduce phase errors into the relationships between the pulses. A calibration on the geometry of the solar sensing assembly is needed. Two types of calibration are possible: 1) geometric calibration by measuring the angles of installation of the solar cell fixtures and deriving a formula relating a to the pulse spacing through the installation angles, and 2) physical calibration using a light source to simulate the sun. Calibration data relating t a n P (2) h(y)=h,+hi + (y -&, ) t anP (3) 2.2 Frictional heat of fluid film The frictional heat flux can be calculated by (4) where is the viscosity of fluid film determined by the mean temperature of fluid film T,. For water, the relationship between its viscosity and temperature[ 151 can be expressed by P( T, ) = Po exp [ -0.0 1 75 ( Tm - To ,] (5) where & is the known viscosity corresponding to the temperature TO. 2.3 Temperature field of sealing rings The heat conductance model of a sealing ring is illustrated in Fig.3, and it is available for both rotating ring and stationary ring. m is the length of the sealing ring and n=Ro-Ri. The heat transfer governing equation is in the form[ 181 a2T a2T ax2 ay2 -+-=o Equation (7) can be obtained by boundary condition at y=n: a -k=cotnk ai Using the thermal insulation conditions at x=O and (8) y=O, the general solution of Eq.(6) was obtained to be m T ( x , y) = Z B , (epl* + e-\"')coskiy i =1 where ki is the solution of Eq.(7) and Bi is undetermined. According to the heat flux boundary condition at x=m, Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001594_bf02575194-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001594_bf02575194-Figure3-1.png", + "caption": "Figure 3. The beating pattern of our two models for symplectie (Opalina) and antipleetic (Paramecium} metachronism. Effective stroke is 2-9 in Opalina, 1-3 in Paramecium", + "texts": [ + " These equations can be solved numerically for a given known movement of a cilium (~) through one cycle of its beat. Unfortunately, little is known about FLOW IN TUBULES DUE TO CILIARY ACTIVITY 519 the movement of cilia in tubes, so the best we can do is to take the movements of those observed in the slower beating cila of protozoa. We will take movement data from OTalina, which exhibits symplectic metachronism, and Paramecium, which has antiplectie metachronism. Their movements are illustrated in Figure 3, and comparisons between the two will be made in the next section. (ii) Interior. At the interface between the cilia sublayer and the interior region, we take the velocity to be continuous, that is, u,(R - L)I,ubtayor = u,(r) 0 < r, lyl < R - L. (12) With this condition we have \"plug\" flow in the interior of the tube. Some velocity profiles for varying ~ , , and ? are shown in Figure 4. A straight velocity profile (\"plug\" flow) need not always be the case. First, we could include in our model the velocity due to a pressure gradient dp/dx along the tube, resulting in the usual parabolic profile defined as: ldp up(y) = 2-~ ~ (y2 _ R2), u,,(r) = 1 gp (r 2 _ R2), 4t, dx two-dimensional model, circular cylinder, 03) where up is the velocity in the x-direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure3-1.png", + "caption": "Fig. 3. A cylindrical body that rotates.", + "texts": [ + " A helical rotation mobile robot does not drag its body because the entire body makes a rolling motion simultaneously. For the same reason, the contact area is wider than for other types of robots, which may dissipate contact force and allow the robot to generate a propulsive force with its entire body. The space inside the helical body offers advantages, because fluid can flow through the space while an inspection is conducted. Since the robot generates force radially, if the tube is deformable and is initially collapsed, the robot can stretch the wall of the tube and then proceed. Figure 3 shows the rotating motion of a cylindrical body that has a two-degree-of-freedom active joint in the middle. The body can bend in any direction and can rotate if one part of the body is fixed. If this body is laid on the ground, its rotating motion generates a rolling motion, as shown in Fig. 4. This is the basic concept of a helical rotation mobile robot. This rolling motion can be achieved by contraction and elongation of wires instead of a two-degree-of-freedom active joint. A pair of wires causes one-degree-of-freedom revolution of a flexible body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003879_j.ymssp.2013.04.009-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003879_j.ymssp.2013.04.009-Figure3-1.png", + "caption": "Fig. 3. Planetary gear complex-valued mode shapes. (a) Planet mode at \u03a9c \u00bc 0:6768 (Point C in Fig. 2), (b) rotational mode at \u03a9c \u00bc 0:5605 (Point F in Fig. 2), and (c) translational mode at \u03a9c \u00bc 0:3905 (point H in Fig. 2). The real and imaginary parts of the complex-valued mode are shown by solid and dashed lines. The gear nominal positions are shown by a dotted line.", + "texts": [ + " For general forcing excitation and initial conditions the response contains contributions from multiple modes, as seen in Eqs. (7) and (5). For such cases, the resulting gear motions will be linear combinations of the single-mode responses described in this section. Planetary gears with equally spaced and diametrically opposed planets have three mode types called planet, rotational, and translational modes [21]. Fig. 2 shows the eigenvalues for each mode type determined from Eq. (2a) (or equivalently Eq. (4)) for the system with four equally spaced planets and parameters in Table 1. Fig. 3 shows example complex-valued mode shapes for each mode type at specific speeds indicated in Fig. 2. The position of each central member (sun gear, ring gear, and carrier) and planet is calculated from the coordinates q\u00f0t\u00de in Eq. (1b) as rh \u00bc x\u0302he1 \u00fe y\u0302he2; h\u00bc c; r; s; \u00f09a\u00de ri \u00bc \u00f0R\u00fe \u03be\u0302 i\u00deei1 \u00fe \u03b7\u0302ie i 2; i\u00bc 1;2;\u2026;N; \u00f09b\u00de where R\u00bc rc=rp is the non-dimensional carrier radius and rp is the planet radius. The response in the stationary basis fE1;E2;E3g is determined using a rotation tensor transformation that relates the two bases according to ek \u00bcQEk for k\u00bc 1;2;3 [37,38]", + " The tensor Q has the matrix components \u00bdQ \u00bc cos \u03a9ct \u2212sin \u03a9ct 0 sin \u03a9ct cos \u03a9ct 0 0 0 1 2 64 3 75: \u00f010\u00de The fei1; ei2; ei3g basis is obtained from the fe1; e2; e3g basis by a rotation of \u03c8 i (the planet position angle) about e3 for i\u00bc 1;2;\u2026;N. The tensor transformation Q i relates the two bases by eik \u00bcQ iek. It has matrix components \u00bdQ i \u00bc cos \u03c8 i \u2212sin \u03c8 i 0 sin \u03c8 i cos \u03c8 i 0 0 0 1 2 64 3 75: \u00f011\u00de Planetary gears have three planet modes where the only non-zero modal displacements are the planet deflections. These modes have natural frequency degeneracy N\u22123 and occur for systems with four or more planets. A representative planet mode is shown in Fig. 3a. Planet modes have the form [21] \u03d5\u00bc \u00bd0;0;0;w1p;\u2026;wNp T ; p\u00bc \u00bd\u03b6; \u03b7;u T ; \u00f012\u00de where wi are, in general, complex-valued when the state eigenvectors from Eq. (4) are chosen orthonormal with respect to A. Alternatively, they can be chosen as real-valued as done in Ref. [21]. The wi are real-valued for four-planet systems with equal planet spacing and have the values w2 \u00bcw4 \u00bc 1 and w1 \u00bcw3 \u00bc\u22121. Use of Eq. (12) in the single-mode free or forced response form in Eq. (8) gives the motion of each planet from Eq", + " In this case the planet motion has multiple loops, with the number of loops equal to the ratio of carrier speed to natural frequency. The curves traced by the planet motions in Figs. 6c and d are similar to lima\u00e7ons. For all cases the motions of each planet in the rotating carrier basis are elliptical orbits, as was the case shown in Fig. 4a at \u03a9c \u00bc 0:5000. Planetary gears have six rotational modes where the central members have only rotation and no translation. Additionally, each planet has identical modal displacements. A representative rotational mode is shown in Fig. 3b. Rotational modes have the form \u03d5\u00bc \u00bdpc;pr ;ps;p1;\u2026;p1 T , where ph \u00bc \u00bd0;0;uh T for h\u00bc c; r; s [21]. For single-mode vibration, each planet has identical motion relative to its local reference frame calculated from Eq. (9b) as ri \u00bc \u00bdR\u00fe 2j\u03b61j cos\u00f0\u03c9t \u00fe \u03b3\u03b6\u00de ei1 \u00fe \u00bd2j\u03b71j cos\u00f0\u03c9t \u00fe \u03b3\u03b7\u00de ei2; \u00f017\u00de where tan\u03b3\u03b6 \u00bc Im\u00f0\u03b61\u00de=Re\u00f0\u03b61\u00de and tan \u03b3\u03b7 \u00bc Im\u00f0\u03b71\u00de=Re\u00f0\u03b71\u00de. An observer rotating with the carrier would measure the oscillation frequency \u03c9. Eq. (17) represents an ellipse, as for the case of single-mode vibration in a planet mode" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001843_1.39537-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001843_1.39537-Figure4-1.png", + "caption": "Fig. 4 Virtual structure rotation.", + "texts": [ + " If the principal axis and the principal angle are known, the coordinate transformation matrix from the inertial frame to the desired frame can be obtained using the following equation [11]: R2 l21 1 cos cos l1l2 1 cos l3 sin l1l3 1 cos l2 sin l1l2 1 cos l3 sin l22 1 cos cos l2l3 1 cos l1 sin l3l1 1 cos l2 sin l3l2 1 cos l1 sin l23 1 cos cos 2 4 3 5 (22) where l l1 l2 l3 T is the principal axis vector, and is the principal angle of the rotation. Because the principal axis, l, is normal to the vectors S and XT, the principal axis vector can be defined as l S XT jS XTj (23) Throughout the virtual structure rotation, each satellite moves to a new position, which can be calculated by using Eq. (22) as A 0 X XA0 X R2XA (24) B 0 X XB0 X R2XB (25) C 0 X XC0 X R2XC (26) Figure 4 shows the qualitative procedure of the virtual structure rotation. Note that S0 is a boresight vector after rotation, and axis l is normal to both S and XT in Fig. 4. C. Third Stage: Targeting The attitude of each satellite is coincided with the VSF in the first and second stages, as shown in Fig. 5. Figure 5 shows the top and side views of the system after the virtual structure rotation. Let us first consider satellite A. The satellite only needs to rotate at the rotationFig. 3 Virtual structure model in 3-D space. D ow nl oa de d by W R IG H T S T A T E U N IV E R SI T Y o n Se pt em be r 11 , 2 01 4 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /1 .3 95 37 angle A about the new principal axis lA to point toward the desired target" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001514_tmag.2009.2018676-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001514_tmag.2009.2018676-Figure4-1.png", + "caption": "Fig. 4. Simulation as using variable alpha angle at beta angle 30 degrees.", + "texts": [ + " Operation characteristic of the spherical wheel motor is estimated with stability of holding torque and accuracy of tilting position. Characteristic of holding torque is related with torque profile which is generated by a coil with a magnet like Fig. 3. The accuracy of tilting position is affected by alpha and beta variation like Figs. 4 and 6. In this paper, therefore, the operation characteristic of the spherical wheel motor is analyzed by using simulation parameters, for instance alpha degree and beta degree of a rotor. Fig. 4 shows simulation as using variable alpha angle at beta angle 30 degrees. Because center positions of upper and lower coils are fixed at alpha 18 degrees, respectively, the stability of holding torque is simulated within alpha 18 degrees. In this simulation, the rotor rotates around the beta angle 30 degrees. Thus a magnet is affected from summed MMF created by 4 coils, for instance upper A, upper B, lower A, and lower B coil. Fig. 5 shows simulation results of rotor position at beta angle 30 degrees" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001463_3.44295-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001463_3.44295-Figure1-1.png", + "caption": "Fig. 1 Component motion of a pneumatic tire.", + "texts": [], + "surrounding_texts": [ + "RELATIONS expressing tire forces in terms of the wheel yaw angle and lateral displacement are required in various studies in aircraft and ground vehicle dynamics, including the analysis of wheel shimmy. There are at present numerous tire theories1\"6 available for calculating these necessary relationships. Unfortunately, from a dynamicist's point of view, these theories are not totally satisfactory. On the one hand, comparatively simple models such as that of Moreland,1 while being of a form convenient for dynamic analysis, suffer from lack of accuracy in characterizing the tire behavior (see Appendix). Conversely, the more accurate theories such as that of Pacejka,6 are complicated to the point of making them extremely cumbersome for direct use in the analysis of dynamical systems. Clearly, an immediate engineering need exists for a practical yet accurate method of representing transient tire forces.13 The development of such a method is the subject of the present paper. The technique to be described here makes direct use of experimental tire data in a frequency response format. This data along with principles from feedback and control systems theory (Bode Analysis) is utilized to develop a set of transfer functions which relate the tire force and moment response to variations in the wheel coordinates. Once these transfer functions are known, a corresponding system of differential equations can be written. It is known that if the tire behaves linearly and the frequency response is represented for all frequencies, then the transient condition is also covered, i.e., the equations are valid for arbitrary motions of the wheel. The differential equations, thus, developed are especially convenient for use in dynamic analysis or simulation studies since they are of the linear constant efficient type. They become part of a set of equations, the rest of which may be nonlinear as necessary to represent the system under analysis." + ] + }, + { + "image_filename": "designv11_3_0002727_1.4023084-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002727_1.4023084-Figure2-1.png", + "caption": "Fig. 2 Equivalent EHL contact model and spring-damper model", + "texts": [], + "surrounding_texts": [ + "The problem to be considered in this paper is the contact vibration between a rolling element and the raceway, as depicted in Fig. 1. The contact pair is under elastohydrodynamic lubrication and the film can be replaced by a spring and a damper, as shown in Figs. 2(a) and 2(b). In point of the mutual approach, it is below zero when the elastic deformation at the contact center is smaller than the central film thickness; otherwise it is above zero. For the oil film force, Ffilm \u00bc \u00f0\u00f0 X pdxdy the Taylor series is as follows: Ffilm \u00bc FS \u00fe @F @h0 h0 h0S\u00f0 \u00de \u00fe @F @ _h0 _h0 \u00fe 1 2 @2F @h2 0 h0 h0S\u00f0 \u00de2 \u00fe 1 2 @2F @ _h2 0 _h2 0 \u00fe @2F @h0@ _h0 h0 h0S\u00f0 \u00de _h0 \u00fe (1) In small oscillations which often occur in mechanical systems, the oil film force can be linearized around the equilibrium position and the two-order and higher-order terms in Eq. (1) can then be neglected. Thus, the present model is linearized as a linear spring and a viscous damper. 2.1 Transient EHL Model. In terms of the dimensionless variables defined in the Nomenclature, the dimensionless Reynolds equation governing the pressure is given by @ @ x e @ p @ x \u00fe 1 j2 @ @ y e @ p @ y \u00bc @ @ x q h\u00f0 \u00de \u00fe @ @s q h\u00f0 \u00de (2) where e \u00bc q h3= gb\u00f0 \u00de and b \u00bc 8pURx= aW0\u00f0 \u00de The boundary conditions are p xa; y; s\u00f0 \u00de \u00bc p xb; y; s\u00f0 \u00de \u00bc p x; ya; s\u00f0 \u00de \u00bc p x; yb; s\u00f0 \u00de \u00bc 0 (3) where p and h represent the hydrodynamic pressure and film thickness, respectively. The dimensionless film thickness at time s is calculated by h x; y; s\u00f0 \u00de \u00bc h0 s\u00f0 \u00de \u00fe x2 2j \u00fe jRx 2Ry y2 \u00fe 2RxpH pE0b \u00f0\u00f0 X p n; k; s\u00f0 \u00deffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi x n\u00f0 \u00de2\u00fej2 y k\u00f0 \u00de2 q dndk (4) where h0 s\u00f0 \u00de denotes the mutual approach and the last item is the surface deformation. The equation of motion for the contact body [3] is as follows: \u20ac h s\u00f0 \u00de=X2 m \u00fe 1:5=p \u00f0\u00f0 X p x; y; s\u00f0 \u00ded xd y \u00bc W s\u00f0 \u00de (5) where Xm \u00bc XnK;Xn \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2Rw0R= mCu2\u00f0 \u00de p ; K \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi RxC= 2RjR\u00f0 \u00de p Xn is the dimensionless natural frequency defined by Wijnant [3]. The initial conditions of Eq. (5) are h0 s\u00bc0 \u00bc h00; _ h0 s\u00bc0 \u00bc _ h00 (6) As the response of the system does not rely on the value of initial deviations in small oscillations [3], the initial conditions of h00 \u00bc 0:9 h0S and _ h00 \u00bc 0 are supposed in the present analysis. The viscosity and density of the lubricant follow the Roelands\u2019 viscosity-pressure equation [13] and Dowson-Higginson\u2019s density-pressure equation [14], respectively. 2.2 Free Vibration Model. Under free vibrations, the load applied on the contact body is assumed to be constant, that is, w\u00bcw0 and W(s)\u00bc 1. With the mutual approach h00 given in the initial moment which deviates from the equilibrium position, the contact body begins to vibrate. The response curve of the mutual approach displacement can be obtained by solving the transient EHL model described in Sec. 2.1 and the curve can be used to get the stiffness and damping of the EHL contact. The method is stated as follows. Replacing the film force by the elastic force and viscous force, the Eq. (5) can be written as \u20ac h0 s\u00f0 \u00de=X2 m \u00fe C _ h0 s\u00f0 \u00de \u00fe K h0 s\u00f0 \u00de h0S\u00bd \u00bc 0 (7) 021501-2 / Vol. 135, APRIL 2013 Transactions of the ASME Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 07/28/2013 Terms of Use: http://asme.org/terms Eq. (7) is rewritten in terms of xn and n as \u20ac h0 s\u00f0 \u00de \u00fe 2nxn _ h0 s\u00f0 \u00de \u00fe x2 n D h0 s\u00f0 \u00de \u00bc 0 (8) where D h0 s\u00f0 \u00de \u00bc h0 s\u00f0 \u00de h0S;xn \u00bc ffiffiffiffiffiffiffiffiffiffi KX2 m q ; n \u00bc CX2 m=\u00f02xn\u00de, representing the displacement of mutual approach, natural frequency and damping ratio of free vibration, respectively. As shown in Fig. 3, the amplitudes of D h0 decay in the proportion of e nxns. Thus, the logarithmic decrement d can be obtained as d \u00bc ln A1=A2\u00f0 \u00de \u00bc nxnsn \u00bc 2pn= ffiffiffiffiffiffiffiffiffiffiffiffiffi 1 n2 q (9) where A1 and A2 are the amplitudes of D h0 and sn denotes the period of oscillation. As the model in this paper is linearized for a nonlinear system, the response curve is not in agreement with the exponentially decay curve well. Thus, the average value of the equivalent logarithmic decrement d in two periods can be more reasonable than its value in one period. As the stiffness and damping coefficients are not time-based, but influenced by the truncation error of the motion equation, the first two periods are ignored and the periods 3 and 4 are used. Using the amplitudes A1 A4, d can be calculated in the following form d \u00bc ln 0:5 A1=A2 \u00fe A3=A4\u00f0 \u00de\u00bd (10) Once n and xn are obtained according to d, the stiffness and damping coefficients can be solved subsequently. All the dimensionless variables used above and the corresponding dimensional forms are defined in the Nomenclature." + ] + }, + { + "image_filename": "designv11_3_0000430_13506501jet215-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000430_13506501jet215-Figure2-1.png", + "caption": "Fig. 2 Schematic diagram of experimental apparatus", + "texts": [ + " TVE \u00bc TVE0 \u00fe A1 ln (1\u00fe A2p) (2) The viscoelastic transition temperature at atmospheric pressure TVE0 was estimated by the occurrence of photo-elasticity effect by lowering the temperature using liquefied nitrogen gas [10]. The parameters obtained are listed in Table 2. The phase diagram of glycerol has been quoted from Herbst et al. [11]. The optical interferometry technique was used to study the shape and thickness of the lubricating film in the EHL point contact. A schematic diagram of the apparatus is shown in Fig. 2. The bodies in contact are a 23.8 mm diameter steel ball (elastic modulus E1 \u00bc 207 GPa, Poisson\u2019s ratio n1 \u00bc 0.3) and a pyrex glass optical flat disc (E2 \u00bc 63.7 GPa, n2 \u00bc 0.2) or a sapphire optical flat disc (E2 \u00bc 365 GPa, n2 \u00bc 0.25) of 45 mm diameter with height of 5 mm. The steel ball was loaded with 86 N via lever arm. The corresponding maximum Hertzian pressure is 0.68 GPa for pyrex glass and 1.35 GPa for sapphire. The optical flat disc is driven by the motor with ring-cone type traction CVT through toothed belt" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001190_icelmach.2010.5608143-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001190_icelmach.2010.5608143-Figure4-1.png", + "caption": "Fig. 4 Distributions of (a) magnetomotive force and (b) permeance.", + "texts": [], + "surrounding_texts": [ + "\u03a6Abstract -- This paper describes the transmission torque characteristics in a surface permanent magnet-type (SPM-type) magnetic gear. The operating principle of this gear and its transmission torque under the synchronous operation in accordance with the gear ratio are formulated. And the highorder components contained in the cogging torque are verified by employing the 3-D finite element method (FEM) and carrying out measurements on a prototype. Furthermore, a method for reducing the cogging torque is discussed.\nIndex Terms--3-D finite element method, cogging torque, magnetic gear, skew.\nI. INTRODUCTION AGNETIC gears have some advantages such as low mechanical loss and maintenance-free operation that are not observed in conventional mechanical gears. In\naddition, they operate as a torque limiter under overloaded condition, and magnetic gears are expected to be applied to a joint of the human robot. However, most of previous magnetic gears have a problem of insufficient transmission torque for practical use due to the narrow facing area between two rotors (Fig. 1).\nRecently, various types of new magnetic gears to solve the above problem were proposed [1]-[5], and a SPM-type magnetic gear employing magnetic harmonics comes to attract attention because of its high transmission torque density though it has a complex structure with multipole magnets as shown in Fig. 2. Some studies on a SPM-type magnetic gear have been carried out, but few papers concerning the cogging torque can be seen.\nThis paper describes the transmission torque characteristics in a SPM-type magnetic gear. The operating principle and the transmission torque under the synchronous operation in accordance with the gear ratio are formulated [6]-[12]. High-order components contained in the cogging torque are computed by employing the 3-D FEM and the result of the analysis is verified by carrying out measurements on a prototype. Furthermore, a method for reducing the cogging torque is discussed.\nII. OPERATING PRINCIPLE Assuming that a low-speed rotor is removed, only a highspeed rotor magnet generates magnetomotive force shown in Fig. 3(a), and stationary pole pieces generate permeance shown in Fig. 3(b), where \u03b8 represents rotor angle.\nIn this model, Fourier series expansions of F(\u03b8) and R(\u03b8) are shown in (1) and (2), respectively.\nN. Niguchi, K. Hirata, M. Muramatsu, and Yuichi Hayakawa are with the Department of Adaptive Machine Systems, Graduate School of Engineering, Osaka University, Yamadaoka, Suita, Osaka, 565-0871 Japan (e-mail:noboru.niguchi@ams.eng.osaka-u.ac.jp).\n{ }\u2211 \u221e\n= \u2212= 1 )12(sin)( m hm NmaF \u03b8\u03b8 (1)\n{ }\u2211 \u221e\n= \u2212+= 1 )12(sin)( l slo NlaRR \u03b8\u03b8 (2) where Nh is the number of pole pairs in the high-speed rotor, Ns is the number of stationary pole pieces. And magnetic flux \u03c6(\u03b8) can be obtained as follows:\n{ }\n{ }[ { } ]\u03b8 \u03b8\n\u03b8\u03b8\u03c6\nhs\nhs l m ml\nm hm\nNmNl\nNmNlaa\nNmRa\n)12()12(cos\n)12()12(cos\n)12(sin)(\n1 1\n1 0\n\u2212+\u2212\u2212\n\u2212\u2212\u2212+\n\u2212=\n\u2211\u2211\n\u2211 \u221e\n=\n\u221e\n=\n\u221e\n=\n(3)\nMagnetic flux contains H1(m), H2(l,m), and H3(l,m) orders shown in (4).\nTransmission Torque Characteristics in a Magnetic Gear\nNoboru Niguchi, Katsuhiro Hirata, Masari Muramatsu, and Yuichi Hayakawa\nM\nXIX International Conference on Electrical Machines - ICEM 2010, Rome\n978-1-4244-4175-4/10/$25.00 \u00a92010 IEEE", + "\u23aa \u23a9\n\u23aa \u23a8\n\u23a7\n\u2212+\u2212= \u2212\u2212\u2212=\n\u2212=\nhs\nhs\nh\nNmNlmlH NmNlmlH\nNmmH\n)12()12(),( )12()12(),(\n)12()(\n3\n2\n1\n(4)\nA high-speed rotor is rotated with \u0394\u03b8, then magnetic flux \u03c6(\u03b8) can be derived as follows:\n{ }\n\u23a5 \u23a6\n\u23a4 \u23ad \u23ac \u23ab \u23a9 \u23a8 \u23a7 \u0394+\u2212\n\u23a2 \u23a3\n\u23a1\n\u23ad \u23ac \u23ab \u23a9 \u23a8 \u23a7 \u0394\u2212+\n\u0394+=\n\u2211\u2211\n\u2211 \u221e\n=\n\u221e\n=\n\u221e\n=\n) ),(\n)()(,(cos\n) ),(\n)()(,(cos\n))((sin)(\n3\n1 3\n2\n1 2\n1 1\n1 10\nmlH mHmlH\nmlH mHmlHaa\nmHRa\nl m ml\nm m\n\u03b8\u03b8\n\u03b8\u03b8\n\u03b8\u03b8\u03b8\u03c6\n(5)\nFrom (5), magnetic flux whose order is H1(m), H2(l,m), and H3(l,m) rotates with \u0394\u03b8, -H1(m)\u0394\u03b8/H2(l,m), and H1(m)\u0394\u03b8/H3(l,m), respectively. To operate as a reduction gear, the number of pole pairs in the low-speed rotor Nl should be equal to H2(l,m) or H3(l,m). Therefore, the relation among Nh, Ns, and Nl can be obtained as follows:\nhls NmNNl )12()12( \u2212\u00b1=\u2212 (6) Then, the gear ratio Gr can be obtained as follows:\nl\nh r N\nNmG )12( \u2212= \u2213 (7)\nIII. ORDERS CONTAINED IN COGGING TORQUE The cogging toque on the high-speed rotor T(\u03b8) under the synchronous operation in accordance with the gear ratio is described in this section. T(\u03b8) consists of the cogging torque generated by stationary pole pieces T1(\u03b8) and by the lowspeed rotor T2(\u03b8).\nAccording to the principle of the transformation from magnetic energy to mechanical energy, equation (8) can be obtained.\n\u03b8 \u03b8\u03b8 \u2202 \u2202\u2212= )()( WT (8)\nwhere W(\u03b8) is magnetic energy. Assuming that magnetic energy is stored only in the air gap, W(\u03b8) is as follows:\n\u222b= V\ndVBW 2 02 1)( \u03bc \u03b8 (9)\nwhere \u03bc0 is the permeability in vacuum, V is the volume of air gap, and B is magnetic flux density in the air gap.\nThe initial rotor angle of the high-speed rotor is represented \u03b8=0, and the rotor angle \u03b4 is given to the highspeed rotor. Then, equation (9) is transformed to (11) with the volume difference of the air gap \u0394V shown in (10).\n\u03b4 \u03c0 \u03b4\u03c0 drlLdrlLV ggsggs 1111 ) 2 2( =\u00d7=\u0394 (10)\n\u222b= \u03b4 \u03bc\n\u03b8 dB rlL W ggs 2\n0\n11\n2 )( (11)\nwhere Ls is the axial length of the air gap, lg1 is the air-gap length between the high-speed rotor and the stationary pole pieces, and rg1 is the average air-gap radius.\nMagnetomotive force and permeance can be defined in Figs. 4(a) and 4(b) by employing \u03b4.\nSubstituting magnetic flux density shown in (12) into (11) yields (13).\n\u03c0 \u03b8\u03b4\u03b4 sg Lr RFB 12 )()( \u2212= (12)\n\u03b4\u03b8\u03b4\u03b4 \u03c0\u03bc\n\u03b8 \u03c0\ndRF Lr\nl W\nsg\ng )()( 8\n)( 2 2\n0\n2 2\n10\n1 \u2212= \u222b (13)\nFourier series expansions of F2(\u03b4) and R2(\u03b4-\u03b8) are shown in (14) and (15).\n\u2211 \u221e\n= += 1 0 2 2cos)( n hnF nNaaF \u03b4\u03b4 (14)\n\u2211 \u221e\n= \u2212+=\u2212 1 0 2 )(sin)( m smR mNbaR \u03b8\u03b4\u03b8\u03b4 (15)\nwhere a0F and a0R are constant. Substituting (14) and (15) into the integral term of (13) yields (16).\n{ }[ { }]\u03b8\u03b4 \u03b8\u03b4\n\u03b8\u03b4\u03b4\n\u03b8\u03b4\u03b4\nssh\nm n ssh\nmn\nm smF n hnR\nRF\nmNmNnN\nmNmNnNba\nmNbanNaa\naaRF\n+\u2212\u2212\n\u2212++\n\u2212++\n=\u2212\n\u2211\u2211\n\u2211\u2211 \u221e\n=\n\u221e\n=\n\u221e\n=\n\u221e\n=\n)2(sin\n)2(sin 2\n)(sin2cos\n)()(\n1 1\n1 0 1 0\n00 22\n(16)\nwhere the first term in right-hand side of (16) becomes zero by the partial derivative by \u03b8, the second, third, and fourth terms also become zero by the integral of \u03b4. Therefore, the constraint that the cogging torque is generated must be that the fifth term is not zero. The cogging torque shown in (17) is obtained as equation (18) is satisfied.\n\u23aa\u23ad\n\u23aa \u23ac \u23ab\n\u23aa\u23a9\n\u23aa \u23a8 \u23a7 \u2212 \u2202 \u2202\u2212= \u222b\u2211\u2211 \u221e\n=\n\u221e\n=\n\u03c0\n\u03b8 \u03c0\u03bc\u03b8\n\u03b8 2\n0 1 1 2 10\n1 1 )sin(\n28 )( m n s\nmn\nsg\ng mNba Lr l T (17)\n02 =\u2212 sh mNnN (18) The cogging torque on the high-speed rotor generated by\nthe low speed-rotor T2(\u03b8) is given by the following procedure as well. Nl-GrNl pole pairs in the low-speed rotor pass the high-speed rotor as the high-speed rotor makes a rotation. Therefore, assuming that the low-speed rotor is fixed, the number of pole pairs in the low-speed rotor can be regarded as Nl-GrNl. The magnetic field by the low-speed rotor is G(\u03b4-\u03b8), and the magnetic field between two rotors F(\u03b4)+ G(\u03b4-\u03b8), is shown as follows:", + "\u2211\n\u2211 \u221e\n=\n\u221e\n=\n\u2212\u2212+\n+=\u2212+\n1\n1 0\n))((cos\ncos)()(\nm lrlm\nn hnFG\nNGNmc\nnNcaGF\n\u03b8\u03b4\n\u03b4\u03b8\u03b4\u03b4 (19)\nwhere a0FG is constant. The permeance between two rotors is constant, R1. The integral term of (13) is represented as follows:\n( )\n))((cos)cos(2\n))((cos2\ncos2\n))((cos\ncos)()(\n1 1\n2 1\n1\n2 10\n1\n2 10\n1\n22 1 2\n1\n22 1 22 1 2 0 2 1 2\n\u03b8\u03b4\u03b4\n\u03b8\u03b4\n\u03b4\n\u03b8\u03b4\n\u03b4\u03b8\u03b4\u03b4\n\u2212\u2212+\n\u2212\u2212+\n+\n\u2212\u2212+\n+=\u2212+\n\u2211\u2211\n\u2211\n\u2211\n\u2211\n\u2211\n\u221e\n=\n\u221e\n=\n\u221e\n=\n\u221e\n=\n\u221e\n=\n\u221e\n=\nlrl m n hnm\nm lrlmFG\nn hnFG\nm lrlm\nn hn\nNGNmnNRcc\nNGNmRca\nnNRca\nNGNmRc\nnNRcRcRGF\n(20)\nwhere only the sixth term generates cogging torque, and it is expanded as follows:\n{ }[\n{ } ]])()(cos[\n])()(cos[ 1 1 2 1\n\u03b8\u03b4\n\u03b8\u03b4\nlrllrlh\nm lrllrlh n nm\nNGNmNGNmnN\nNGNmNGNmnNccR\n\u2212+\u2212\u2212+\n\u2212\u2212\u2212+\u2211\u2211 \u221e\n=\n\u221e = (21)\nCogging torque is generated as equation (22) is satisfied. The cogging torque generated by the low-speed rotor T2(\u03b8) is represented in (23).\n0)( =\u2212\u2212 lrlh NGNmnN (22)\n\u23aa\u23ad\n\u23aa \u23ac \u23ab\n\u23aa\u23a9\n\u23aa \u23a8 \u23a7 \u2212 \u2202 \u2202\u2212= \u222b\u2211\u2211 \u221e\n=\n\u221e\n=\n\u03c0\n\u03b8 \u03c0\u03bc\u03b8\n\u03b8 2\n0 1 1\n2 12\n20\n2\n2\n)(cos 8\n)(\nm lrl n nm sg\ng NGNmccR Lr\nl\nT (23)\nwhere lg2 is the air-gap length between rotors, and rg2 is the average air-gap radius.\nThe cogging torque in the high-speed rotor T(\u03b8) is represented as follows:\n\u23aa\u23ad\n\u23aa \u23ac \u23ab \u2212\n\u23aa\u23a9\n\u23aa \u23a8 \u23a7 \u2212 \u2202 \u2202\u2212=\n\u222b\u2211\u2211\n\u222b\u2211\u2211 \u221e\n=\n\u221e\n=\n\u221e\n=\n\u221e\n=\n\u03c0\n\u03c0\n\u03b8 \u03c0\u03bc\n\u03b8 \u03c0\u03bc\u03b8\n\u03b8\n2\n0 1 1 2 10\n1\n2\n0 1 1\n2 12\n20\n2\n)sin( 28\n)(cos 8\n)(\nm n s\nmn\nsg\ng\nm lrl n nm sg\ng\nmNba Lr l\nNGNmccR Lr\nl\nT (24)\nThe cogging toque generated in the low-speed rotor can be obtained as well. Therefore, the order contained in cogging torque is summarized in Table I.\nIV. FINITE ELEMENT MODEL\nTo verify the orders contained in cogging torque, 3-D FEM is employed. To solve the Maxwell's equation, the magnetic vector potential A is employed, and in this study, the eddy current is ignored. Then the fundamental equation is as follows:\nM\u0391 rot)rot(rot 0\u03bd\u03bd = (25) where v and v0 are the reluctivity of magnetic material and vacuum, respectively, M is the magnetization. Transmission torque is calculated by the nodal force method.\nThe gear ratio of the analysis model is decided in accordance with (6). In this study, the specifications are shown in Table II, and the dimensions are shown in Table III. The analysis model is shown in Fig. 5. In this model, 14 pole pieces are connected by the flux path facing to the highspeed rotor, whose width is 0.5 mm, equal to the thickness of the laminated silicon steel sheet. This structure helps not only to assemble, but also reduce the cogging torque. However, it is thought that the maximum transmission torque slightly decreases due to the short-circuit magnetic flux.\nV. TRANSMISSION TORQUE ANALYSIS\nA. Analysis Condition" + ] + }, + { + "image_filename": "designv11_3_0002564_j.asoc.2011.03.030-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002564_j.asoc.2011.03.030-Figure2-1.png", + "caption": "Fig. 2. The experimental system: (a) the test rig; (b) schematic of the two-stage gearbox.", + "texts": [ + " For very time\u2013frequency matrix A, the feature matrix F obtained by ur TD-2DLDA can be calculated by: = VT AW (13) he dimension of the feature matrix F is g \u00d7 d, which is much less han the dimension of feature matrix Y obtained by the original DLDA. Furthermore, both the row and column information of the riginal matrix A is considered by the TD-2DLDA scheme. In the cost ing 11 (2011) 5299\u20135305 5301 of calculate the 2DLDA algorithm twice, the TD-2DLDA can obtain a more compact feature matrix that preserve more information of the original matrix than 2DLDA. 4. Application of TD-2DLDA to gear fault diagnosis 4.1. Gear experiment system The vibration data of gear used in this paper was acquired from a two stage gearbox, as shown in Fig. 2. Single point faults were introduced to the test gears. Two different type fault types, i.e. gear root crack and tooth wear were set to middle shaft gear B and output shaft gear D respectively in the experiment. Therefore, five gear states i.e. normal condition, middle shaft gear tooth crack, middle shaft gear tooth wear, output shaft gear tooth crack and output shaft gear tooth wear were simulated in the test for evaluating the proposed methods. Six accelerator sensors, as shown in Fig. 2(b), were attached on the six bearing bases to pick up the vibration signals. In the experiment, the motor speed was set to be 1330 rpm and the load was set to be 100 NM. In this work, only the signals acquired from sensor S6 are analyzed. The sample frequency was set to be 4096 Hz and the sample point was 4096 in the experiment. Forty samples were collected for each state, thus totally 200 samples were collected. For each sample of the gear, a 2048 \u00d7 4096 time\u2013frequency matrix can be obtained based on the S transform" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003479_icra.2012.6224994-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003479_icra.2012.6224994-Figure2-1.png", + "caption": "Fig. 2. Architecture of the ball catching system.", + "texts": [ + " In order to complete the terms required in (6), the two (3\u00d7 1) vectors \u02c6\u0307pc o and \u03c9\u0302 c o should be computed. The former can be again retrieved by the previous numerical integration, while the latter can be obtained as \u03c9\u0302 c o = (p\u0302c o \u00d7 \u02c6\u0307pc o)/\u2016p\u0302 c o\u2016 2. Since before of nl measurements it is not possible to have any estimation of p 0 and p\u0307 0 , an initial rough estimation should be provided in order to compute the above quantities. For such a reason, a statistical calibration has been preliminary realized, and the results have in turn been employed in the experiments presented in the next section. Figure 2 shows the experimental set-up implementing the proposed control algorithm. A Comau Smart-Six robot manipulator mounted on a sliding track and equipped with a 4-fingered hand composed of 16 Bioloid Dynamixel AX-12 servomotors has been employed. The Comau C4G control unit is in charge of the compensation of the robot dynamic model, while an external PC with Ubuntu OS patched with the RTAI-real time kernel generates the position/orientation references at 2 ms. The control PC communicates with a second Windows OS PC that is responsible of the visual elaboration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002271_s00216-012-6598-y-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002271_s00216-012-6598-y-Figure4-1.png", + "caption": "Fig. 4 Schematic of OLGA manifold", + "texts": [ + " 3 between 0 and 500 \u03bcM showed the custom-built laminar flow cell to have an extended linear range (R200.995), compared to the wall-jet flow-through cell (R200.936). Subsequently, the rest of our trials were performed exclusively with the custom-built laminar flow cell. Lactosenz TM\u2014prototype design Meeting the analytical needs of the bioprocessing, environmental and industrial sectors requires sensor devices to combine the necessary performance characteristics with liquid handling systems that can automatically regulate incoming samples to be within optimal limits for the sensor. Figure 4 depicts the measurement cycle, beginning with sample acquisition from a dairy processing wastewater drain; the sample was continuously removed at 3.6 mL min\u22121 from the drain by a peristaltic pump. Particles larger than 0.2 \u03bcm were removed by a cross-flow nitrocellulose filter disc (OLAF), and the diffusate from the filter was pumped at 1.3 mLmin\u22121 into a bubble trap to ensure the sample stream was purged of particulates and bubbles prior to delivery to the auto-dilution device (OLGA). The injection manifold of OLGA comprises four valves working in tandem with a step-motor-driven high-precision pump" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure2-1.png", + "caption": "Fig. 2 Geometry of the aluminum prototype wheel", + "texts": [ + " The final component is an inner cover plate that covers the impeller and is equipped with an inlet for the coolant supply (Fig. 1d). In the following chapter the design process of the GIC will be described. 3.2 Adaption of the geometry The design of the developed GIC is based on the design of the aluminum prototype wheel presented in [13]. The aluminum prototype wheel, with a total width of 30 mm, provides channels with a height of 10 mm. The outlets of the cooling channels cover 23% of the complete circumferential surface and 70% of the circumferential surface related to the height of the channels respectively (see Fig. 2). Therefore, only 30% of the circumferential surface can be used for the abrasive layer (related to the height of the channels). In the area of the emersions, with only 30% of abrasive layer, high wheel wear can be expected, resulting in low tool life. To increase the tool life of the GIC, the area for the abrasive layer had to be enlarged. Therefore, the geometry of the outlets had to be adapted. It was necessary to create a transition from the channel geometry inside the grinding wheel (inner geometry) to the geometry at the circumferential of the grinding wheel (outer geometry) (see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000762_ac8024619-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000762_ac8024619-Figure1-1.png", + "caption": "Figure 1. Fabrication of polymeric particles via a solvent displacement method.", + "texts": [ + " The mole fraction of the protonated form of the chromoionophore is related to the fluorescence signal as 1-R) [IndH+] IndT ) 1+ Fmax -F Fmax -Fmin (2) where F is a fluorescence intensity ratio (at two wavelengths) measured in a given experiment, Fmin, and Fmax are the fluores- cence intensity ratios at the minimum and maximum protonation of the chromoionophore, respectively, [IndH+] is the concentration of the protonated form of the chromoionophore. The intensity ratios at the minimum and maximum protonation were measured in 0.01 M solutions of NaOH and HCl, respectively. The pH of the solutions was maintained with 10 mmol of TRIS and MOPS buffers. All experiments were conducted at ambient temperature (23 \u00b1 2 \u00b0C). Activity coefficients were calculated according to Debye-Hu\u0308ckel formalism.21 The solvent displacement method19,20 involves two essential steps (Figure 1). During the first step, a preformed hydrophobic polymer is dissolved in a solvent miscible with water. Injection of a polymer solution into a stirred aqueous phase causes spontaneous emulsification. Polymer deposition at the interface between the organic (disperse) and aqueous (continuous) phases instantaneously produces a colloidal suspension. A surfactant is usually added to the aqueous phase in order to stabilize the emulsion. The solvent diffuses into the aqueous phase causing the micro or nanospheres to precipitate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001564_ecce.2009.5316091-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001564_ecce.2009.5316091-Figure1-1.png", + "caption": "Fig. 1. Principle construction of a ball bearing with the required geometrical", + "texts": [], + "surrounding_texts": [ + "The procedure for obtaining the frequency response of the mechanical system of a drive was explained in detail in earlier papers [20]. The characteristic fault frequencies produced by the bearings that can be seen in the frequency response can be given as: b ORF n c dz f f 1 cos( ) 2 d = \u22c5 \u22c5 \u2212 \u22c5 \u03b8 \u239b \u239e \u239c \u239f \u239d \u23a0 (1) and b IRF n c dz f f 1 cos( ) 2 d = \u22c5 \u22c5 + \u22c5 \u03b8 \u239b \u239e \u239c \u239f \u239d \u23a0 In case the required mechanical bearing data are not available, the following approximation for determining the characteristic fault frequencies can be utilized for 8 \u2264 z \u2264 12. \u03c0 \u03a9 \u22c5\u22c5= 2 z4,0f~ n ORF (2) and \u03c0 \u03a9 \u22c5\u22c5= 2 z6,0f~ n IRF variables The bearing damages produce a periodic disturbance torque as well as a disturbance in the velocity signal [19], [21], [22]. These disturbances lead to changes in the frequency response of the mechanical system at the frequencies fS,ORF and fS,IRF respectively [23]. For calculating the frequency response of the mechanics in the closed loop speed control, the torque generating component of the stator current iq(t) and the velocity of the motor nM(t) are utilized. For that reason the frequency response analysis yields more reliable results than the frequently used one channel FFTanalysis. The frequency response of motor under no load conditions is given by: ( ) ( ) ( ) Mq M mech Js 1 si ssG \u22c5 =\u03c9= (3) The frequency response Gmech(j\u03c9) of the investigated two- inertia-system is given by: ( ) ( ) ( ) ( ) 2T T j dT j 11 L C CG jmech T T TT T j 2M L CM L j dT j 1CT TM L \u03c9 + \u22c5 \u03c9+ \u03c9 = \u22c5 + \u03c9 \u03c9 + \u22c5 \u03c9+ + (4) The calculation of the frequency response is carried out by applying the Welch-method, as explained in earlier papers [20]. As the measurement is carried out in a closed loop control system, two measurement signals are required in order to get reliable results. In the case under consideration the frequency response of the mechanical system Gmech(j\u03c9) results from the following relationship in which the hat symbol denotes estimated values: ( ) W W\u02c6 \u02c6P P, ,M M M MG jmech W W\u02c6 \u02c6P P,i ,M q M M \u0394\u03c9 \u03c9 \u0394\u03c9 \u0394\u03c9 \u03c9 = \u0394\u03c9 \u0394\u03c9 \u0394\u03c9 (5) The measurement is carried out in closed loop speed control. The system is excited by PRBS, which is fed in at the output of the PI-speed controller. Fig. 2 displays the structure of the setup in case of a one-inertia-system and Fig. 3 shows the measurement configuration with a mechanical twoinertia-system. In both cases the closed loop current control loop is approximated as first order lag." + ] + }, + { + "image_filename": "designv11_3_0003347_j.proeng.2012.09.530-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003347_j.proeng.2012.09.530-Figure1-1.png", + "caption": "Fig. 1. The motion of continua", + "texts": [ + ": +36-52-415-155; E-mail address: tamas.mankovits@eng.unideb.hu \u00a9 2012 Published by Elsevier Ltd.Selection and/or peer-review under responsibility of the Branch Offi ce of Slovak Metallurgical Society at Faculty of Metallurgy and Faculty of Mechanical Engineering, Technical University of Ko\u0161ice Using the Lagrangian description in the reference cylindrical coordinate system the motion of continua ),( 0 trrr = , in the current configuration at time t is given by the reference coordinates 0r of the material points, see in Fig. 1., where u is the displacement. In the subsequent formulation the deformation gradient 0r r F \u2202 \u2202= (1) is multiplicatively decomposed into a volumetric changing part V F and a volume preserving part F\u0302 FFF V \u02c6= (2) where IJF 3 1 \u02c6 \u2212 = , 1\u02c6det =F (3) and FJF V 3 1 = , JFF V == detdet (4) where J is the Jacobian. Using this decomposition we can define the so called right Cauchy-Green tensor and its volumetric preserving part FFC T= , CJFFC T \u02c6\u02c6\u02c6\u02c6 3 2\u2212 == (5) where T denotes the transpose of a tensor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000964_j.eswa.2009.03.027-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000964_j.eswa.2009.03.027-Figure1-1.png", + "caption": "Fig. 1. Overall setup of the rotary inverted pendulum system: (a) experimental setup and (b) details of the underlying mathematical description of the system.", + "texts": [ + " The rotary inverted pendulum system is composed of a pendulum, rotating arm, potentiometer, and a motor. The pendulum system is unstable: a motion of the pendulum becomes unstable when pendulum is placed in a vertical position (Musknja & Tovornik, 2006). In this control system, the rotating arm is revolved with the use of the motor which controls the motion of pendulum. The objective of the controller is to control position of the rotating arm and to stabilize the pendulum in its inverted upright position. The overview of the system is presented in Fig. 1. We define the positive direction of the rotating arm to be counterclockwise; likewise the positive direction of the pendulum is also clockwise. The pendulum can be described as a system with a single input Vm (voltage of the motor), and two outputs: a (angle of the pendulum) and h (angle of the rotating arm). The motion equations of the rotary inverted pendulum are those of Euler\u2013 Lagrange which govern a total potential and kinetic energy. The detailed nonlinear motion equations are as follows: \u00f0Jeq \u00femr2\u00de\u20ach mLr cos\u00f0a\u00de\u20aca\u00femLr sin\u00f0a\u00de _a2 \u00bc Toutput Beq _h \u00f01\u00de 4 3 mL2 \u20aca mLr cos\u00f0a\u00de\u20ach mgL sin\u00f0a\u00de \u00bc 0 \u00f02\u00de Toutput \u00bc gmggktkg\u00f0Vm kgkm _h\u00de Rm \u00f03\u00de The linearized version of these equations can be obtained by the use of the commonly accepted substitutions made in (1)\u2013(3), that is cos\u00f0a\u00de \u00bc 1 and sin\u00f0a\u00de \u00bc a The values of the parameters of the system are summarized in Table 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002918_2012-36-0305-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002918_2012-36-0305-Figure3-1.png", + "caption": "Figure 3. FZG gear test rig: (a) Picture and (b) schematic view.", + "texts": [ + " All equipment used in mechanical testing belongs to the Metallurgical and Mechanical Laboratory of Tupy SA. Table 4 summarizes the results obtained after mechanical tests. Table 4 \u2013 Mechanical properties of spur gears materials. Material Tensile Strength (MPa) Elongation (%) Young Modulus (GPa) Hardness (HRc) Charpy Energy (Joules) AISI 8620 1325* 12* 205* 58 -- AISI 4140 2040* 7,6* 205* 56 4,5 ADI 1 1546 2,1 -- 56 -- ADI 2 1273 3,5 186,9 54 5,8 * (Spur gears materials). For the contact fatigue tests was used a FZG-LASC tribometer. Figure 3 shows an overview of this equipment. This tribometer was designed, manufactured and assembled in the Contact and Surface Laboratory (LASC-UTFPR). By using the power recirculation principle, two pairs of gears can be tested at the same time. The load is imposed on the gears by applying torque on the shaft that the wheel is mounted on (FZG loads k6 and k9). A twist on the wheel axis is achieved by applying an eccentric load, using a lever and dead weight. To produce accelerated wear on the flank of gear teeth, it is common to use gears with modified profile" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003347_j.proeng.2012.09.530-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003347_j.proeng.2012.09.530-Figure4-1.png", + "caption": "Fig. 4. The air-spring", + "texts": [ + " The volumetric change and the hydrostatic pressure are approximated by lower order of polynomials than the displacement zaraapJ 210)2( ++== , (20) zbrbbpp 210)2( ++== . (21) After the discretization we obtain the formula for the Newton-Raphson iteration fffuK \u0394=\u2212=\u0394 intextT , (22) where T K is the tangential stiffness matrix, u\u0394 is the displacement increment and f\u0394 is the unbalanced load vector. A p-versional finite element code has been developed for the analysis of nearly incompressible materials. The approximation order for the displacement is 2=p . A rubber bumper in an air-spring (see in Fig. 4) is analyzed by the FEM code based on the theory discussed above. The load-displacement curve for pressure is known from the producer but only when the bumper is loaded between two flat rigid surfaces. A parameter optimization is needed to find the material parameters. The measurement and the code were compared, see in Fig. 5. A numerical analysis is also needed to obtain the optimal input parameters for the finite element analysis such as mesh density and the so called penalty parameter. The effect of the mesh density change was compared at two discrete points (7 and 14 mm compression)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003294_jjap.51.06fl17-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003294_jjap.51.06fl17-Figure7-1.png", + "caption": "Fig. 7. (Color online) Fabrication of a metalized microscrew. (a) Polymer model of a microscrew produced by two-photon microfabrication. The scale bar is 10 m. (b) Metalized microscrew before laser ablation. (c) Released microscrew after laser ablation. (d) Rotation of the released microscrew in water.", + "texts": [ + " 5(d). The three anchors attached to the blades were sequentially removed one after the other by 06FL17-3 # 2012 The Japan Society of Applied Physics focusing the femtosecond pulsed laser beam. To demonstrate the rotation of the microrotor, the released microrotor was immersed in water, and then it was rotated by a glass needle as shown in Fig. 6. It was confirmed that the undamaged metalized microrotor rotated. We also succeeded in fabricating more sophisticated 3D metallic movable microparts. Figure 7(a) shows an SEM image of a metalized microscrew model with bearings. The rotating screw was initially supported with two anchors [Fig. 7(b)]. After laser ablation, the anchors were completely removed without any damage to the screw [Fig. 7(c)]. The released screw was also rotated by a glass needle in water [Fig. 7(d)]. These results demonstrate that the combination of the anchor supporting method and release by laser ablation is useful for making metallic 3D movable microparts with complex shapes. We proposed a fabrication process for 3D metalized movable microparts by using two-photon microfabrication and electroless plating. In our method, the electroless plating of 3D polymer models produced by two-photon microfabrication can provide high-precision, sophisticated 3D metalized microstructures. In addition, the anchors used to support metalized movable microparts were easily removed by laser ablation using a femtosecond pulsed laser beam" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001338_j.jmapro.2009.07.002-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001338_j.jmapro.2009.07.002-Figure2-1.png", + "caption": "Fig. 2. Assembly of the electromagnetic vibrator.", + "texts": [ + " To investigate the in-process mechanical vibration on the quality of laser cladding, a vibration system is integrated into the laser-cladding system, which consists of : (1) a 1-kWNd:YAG laser system with a laser head capable to deliver the powder, (2) a 5- axis CNC vertical machining center with a laser head attached to the Z-axis, (3) a powder feeding system capable of delivering metal powder with the protective argon gas flow, and (4) the vibrator setup, which is placed on the x\u2013y-table of the CNC vertical machining center. The schematic of the vibrator system is shown in Fig. 1. An electromagnetic shaker [20] is tightly connected to an aluminum plate (shown in Fig. 2) that is placed on the x\u2013y table of the CNC machine. The frequency of the vibration is controlled by a signal generator that is connected to a 400-W amplifier. The shaker is equipped with an accelerometer that is connected to a data acquisition system to monitor the vibrator parameters. shed by Elsevier Ltd. All rights reserved. A CNC program is developed to define the deposition path and the process parameters. Based on the previous experiences in deposition of AISI H13 powder with MultiFab1 system, the process parameters are chosen in a way that an optimum deposition in terms of clad geometry and deposition rate is obtained" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003725_icuas.2015.7152366-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003725_icuas.2015.7152366-Figure3-1.png", + "caption": "Fig. 3. MAVion free-body diagram: aerodynamic/propulsion forces and moments conventions and directions.", + "texts": [ + " Notice that is the case because the motors are installed in a counter-rotative tandem configuration with rotation direction chosen to counter wing tip vortices, which artificially increases the aspect ratio. It also provides a natural way to trigger banked turns since increasing \u03c91 would not only result in a positive yaw moment but also in an induced positive rolling moment yielding a starboard curve in horizontal flight. Finally, lift is generated by the aerodynamic shape of the fuselage where inside resides most electronic parts and payload. Figure 3 illustrates the MAVion free-body diagram during an arbitrary flight instant. We assume symmetry in counterrotating propellers speed \u03c9 and elevons deflection \u03b4 to cancel out lateral/directional dynamics and focus on longitudinal flight. Under these assumptions, convenient system state x\u2032, actuator u and disturbance w definitions are x\u2032 = vn vd \u03b8 \u03b8\u0307 , u = ( \u03c9 \u03b4 ) , w = ( wn wd ) (1) where vn, vd, wn, wd and \u03b8, denote, respectively, MAVion and wind velocities with respect to the ground expressed in the local north-east-down (NED) coordinate system, and pitch angle with respect to the local horizontal. In order to calculate x\u0307\u2032, we apply Newton\u2019s second law (see Fig. 3) and obtain mv\u0307n = (D + 2T )cos\u03b8 + Lsin\u03b8 (2) mv\u0307d = mg + Lcos\u03b8 \u2212 (2T +D)sin\u03b8 (3) Jy \u03b8\u0308 = M (4) where g, m and Jy denote, respectively, local gravity, MAVion mass and inertia moment with respect to the y-axis. Notice that, otherwise stated, aerodynamic drag D and lift L, aerodynamic pitching moment M , and propulsion thrust T are hereafter described in body-fixed axes. By means of the Buckingham-\u03a0 theorem [8] and assuming inviscid incompressible flow, we relate the aerodynamic forces and moments to the system state according to L = 1 2 \u03c1v2\u221eSCL(\u03b1, \u03b4) (5) D = 1 2 \u03c1v2\u221eSCD(\u03b1, \u03b4) (6) M = 1 2 \u03c1v2\u221ec\u0304SCM (\u03b1, \u03b4) (7) where \u03c1, v\u221e, S, c\u0304, \u03b1, CL, CD and CM are, respectively, air density at drone location, true air speed with respect to the drone, reference wing area and chord, angle of attack; and lift, drag and pitching moment coefficients" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002263_pime_conf_1969_184_427_02-Figure9.2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002263_pime_conf_1969_184_427_02-Figure9.2-1.png", + "caption": "Fig. 9.2", + "texts": [ + "10) Proc lnstn Mech Engrs 1969-70 Vol184 Pt 3N at Gazi University on April 12, 2016pcp.sagepub.comDownloaded from c( may be complex, its phase denoting the angular difference between r and F. If the journal member of a seal has a steady precession speed wp at radius r, then, in equation (9.6), it is possible to put Y = r &mpt, with F = -d(FU2+F2) = - jr ejcIpt it is then found that ,e- j@([c~,- l /Zws]T) (9.11) h 1 Q = -. TRP S ( [ W ~ - + W ~ ] T ) The significance of equations (9.10) and (9.11) is indicated in Fig. 9.2. Two important special cases will be mentioned : The analyses described so far are linearized, assuming small journal eccentricity ratio. The extent to which these results may be applied to larger eccentricities can be judged Proc lnstn Mech Engrs 1969-70 Vol184 Pt 3N at Gazi University on April 12, 2016pcp.sagepub.comDownloaded from 0 4 6 8 10 TTRP A a- I oo 2 oo 30' only on the basis of a more complete calculation, free from linearization. This is too complex to admit formal analytical solution, but is amenable to direct digital computation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003171_0954406211399659-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003171_0954406211399659-Figure1-1.png", + "caption": "Fig. 1 Cross-sectional view of internal gear pump", + "texts": [ + " Wilson proposed the mathematical model of the friction torque for various hydraulic pumps and motors [10,11]. However, it was constructed conceptually and was not suitable for the balanced vane pump and the internal gear pump. Although the author also proposed another mathematical model of the friction torque for the balanced vane pump, it cannot be applied to the internal gear pump. Therefore, based on the previously proposed models, a new mathematical model to represent the friction torque in the internal gear pump is deduced from the experimental results of the friction torque. Figure 1 shows a cross-sectional view of an internal gear pump used for transmission of AT or CVT, composed of an inner rotor with external gear teeth, an outer rotor with internal gear teeth, a body holding both the rotors, a cover, and a stator shaft supporting the inner rotor via a bush bearing. The stator shaft is fixed to the cover and does not rotate. The bush bearing is inserted in the inner rotor and it makes the rotation of the inner rotor on the stator shaft smooth without seizure under high-pressure conditions", + " In comparison to the gerotor pump, the internal gear pump with this tooth profile has a larger displacement volume per gear tooth cavity and the radial dimensions of both the rotors in this pump can be smaller than those in the gerotor pump when the pump displacement and the width of both the rotors are the same. The reduction in the radial dimensions of the rotors in the internal gear pump leads to an improvement in the mechanical efficiency with a decrease in the viscous friction caused by shearing oil at the sides of the rotors. In this study, two internal gear pumps are tested. First, the dimensions of pump no. 1 are given in Fig. 2. In the transmission of AT or CVT, the inner rotor in the pump shown in Fig. 1 is directly driven by input shaft of the transmission, which is rotated by the driving force from a crankshaft of an engine. In this experiment also, the inner rotor was directly driven by a driving shaft separated from the test pump. The driving shaft was rotated by an electric motor and a torque meter was equipped between the electric motor and the driving shaft. The inlet and outlet pressures were measured at the inlet and outlet of the pump, respectively. The pressure differential across Proc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002321_jps.2600561105-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002321_jps.2600561105-Figure7-1.png", + "caption": "Fig. 7-Standard calibration curve. Key: 0, TC; A, DMCTC; n, CTC.", + "texts": [ + " Figures 4-6 show the effect of alkaline methanol, an auxochromic agent, on the absorption spectraof DMCTC, TC, and CTC. In each instance the alkaline methanol had a hypochromic as well as a hyperchromic effect. The wavelengths where maximum absorption occurred when alkaline methanol was introduced into the tetracycline chromaphoric system are as follows. DMCTC, 372 mp; TC, 366 mp; CTC, 377 mp. Standard calibration curves of DMCTC, TC, and CTC hydrochlorides were prepared with alkaline methanol reagent in chloroform solutions. Three standard curves are presented in Fig. 7. Typical linear Beer's law plot of absorbance versus tetra- 1399 1.0 W 0 z L1: 2 8 0.5 m U Fig. 5-A SO CHaOH- CH ( '. - 300 400 5 WAVELENGTH, rnp ion spectra of TCH.Cl. Key: - - - -, I; -, NH~OH-CH~OH-CHC~S. \\ \\ I I ' I 300 400 500 WAVELENGTH, rnp Fig. 6-Absorption spectra of CTCdHCl. Key: _ _ - - , CHsOH-CHCl3; ~ NH4OH-CH8OHCHCE3. cycline concentration range from 0.1 mg./50 ml. to 1.0 mg./50 ml. for all three tetracyclines. The absorptiometric system obtained a maximum absorbance value immediately upon the addition of the alkaline methanol reagent and this absorbance remained constant over a practical length of time" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002425_hri.2010.5453183-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002425_hri.2010.5453183-Figure1-1.png", + "caption": "Figure 1. Markers and angles used to describe head motions.", + "texts": [ + " The distance between the subjects was set as large as possible while allowing the motion capturing of both subjects, resulting in about 1 m of separation. Directional microphones (Sanken CS1) were positioned pointing towards each subject. The motion capture system used is the Hawk system from Motion Analysis. Ten infra-red cameras are arranged in a rectangle around a room in order to capture the motions of both speakers. Seven hemispherical passive reflective markers are applied to the speaker\u2019s head, nose and chin, as shown in Fig. 1. The markers on the head and nose provide a static reference frame for the (rigid) head, while the marker on the chin (relative to the nose marker) was used to align the motion data with the speech audio data, since systematic errors sometimes occurred in the synchronization between these data. It is worth mentioning that although the situation with motion capture markers would be experienced as unnatural, all speakers agreed that this unnatural feeling happened only in the beginning of the first time they experienced the motion capture markers. After a while (say, 1 or 2 minutes), they simply forget about the markers, and can talk naturally. The three rotation angles, shown in Fig. 1, are used to describe the head motions. We use the terms \u201cnod\u201d, \u201cshake\u201d and \u201ctilt\u201d, in correspondence with the terms \u201cpitch\u201d, \u201cyaw\u201d, and \u201croll\u201d, used in aerodynamics. The head rotation angles are estimated from the markers based on the singular value decomposition method [12], given by the following expressions. [U,D,VT] = svd ( reference * target ), (1) where reference and target are the 3D marker set of the neutral and current positions respectively, translated to new coordinates having their centroids as origin", + " In a second model (MODEL II), the nod timing generation of MODEL I was superimposed by a slight face up motion (3 degrees) during the speech utterance intervals. This upward motion during speech was also often observed in natural speech, and is expected to improve the motion naturalness with regard to using only single nods as in MODEL I. Six conversation passages including nods with relative occurrence were randomly selected from our database, and rotation angles (nod, shake and tilt angles, as in Fig. 1) were extracted for each utterance. The duration of the conversation passages was limited to 10 to 20 seconds, since subjects have to compare a pair of motions for the same speech utterances, and such comparison would be difficult if each stimulus is too long. Also, as a dialogue act is attributed for each phrase unit, utterances of 10 to 20 seconds usually contain more than 10 phrases, i.e., some context information is still present in the dialogue passage. For each conversation passage, two types of motion were generated, one for each of the nodding generation models described in the previous sub-section (MODEL I, MODEL II)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003189_j.cnsns.2011.04.012-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003189_j.cnsns.2011.04.012-Figure1-1.png", + "caption": "Fig. 1. (a) Model of a rigid rotor supported by two aerodynamic journal bearings and (b) cross section of a four-lobe bearing.", + "texts": [ + " Therefore, finite element methods have been employed to obtain the solution and then Runge\u2013Kutta method has been used to solve this equation and equations of motion simultaneously to obtain position, velocity and acceleration of the rotor center. d Method x y s = 10 s = 100 s = 1000 s = 10 s = 100 s = 1000 0.4 FEM 0.029450 0.007134 0.034625 0.015963 0.032084 0.001221 FDM 0.029190 0.006686 0.034124 0.015174 0.031490 0.001588 0.5 FEM 0.035754 0.068212 0.020073 0.064393 0.058021 0.047436 FDM 0.033392 0.067159 0.021221 0.063349 0.058758 0.045936 The geometric details of a four-lobe noncircular bearing configuration are shown in Fig. 1. Analysis of aerodynamic four-lobe bearing involves solution of the governing equations separately for an individual lobe of the bearing, treating each lobe as an independent partial bearing. To generalize the analysis for all noncircular geometries, the film geometry of each lobe is described with reference to bearing fixed Cartesian axes (Fig. 1). Thus, the film thickness in the clearance space of the kth lobe, with the rotor in a dynamical state is expressed as h \u00bc C \u00f0Xj\u00de cos h \u00f0Yj\u00de sin h\u00fe \u00f0C Cm\u00de cos h hk 0 \u00f01\u00de where \u00f0Xj;Yj\u00de is the rotor center coordinate in the dynamical state and hk 0 is angle of lobe line of centers. C and Cm are conventional radial and minor clearances, when journal and bearing geometric centers are coincident. The pressure governing equation of isothermal flow field in a bearing lobe is modeled by the Reynolds equation as follow [11] @ @X h3P @P @X ( ) \u00fe @ @Y h3P @P @Y ( ) \u00bc 6 l U @ @X \u00fe 2 @ @ t \u00f0P h\u00de \u00f02\u00de in which P is the absolute gas pressure, l is the gas viscosity, U is the peripheral speed of the rotor and t is the time" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003151_ecce.2013.6647071-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003151_ecce.2013.6647071-Figure3-1.png", + "caption": "Figure 3. Static rotor eccentricity of PM motor.", + "texts": [ + " Six current probes are used to measure the current in the coils, i.e., U1+, U2+, V1+, V2+, W1+, and W2+ at the same time. These outputs are then saved in a data logger. The operating conditions are under no load operation at a frequency of 50 Hz. Figs. 1 and 2 are diagrams of stator winding equivalent circuits: Fig. 1 shows the circuit with the connected neutral point (NP connected) while Fig. 2 shows that with the unconnected (NP non-connected). Both types use the same motor, which can be switched to neutral point on or off. Fig. 3 outlines the static rotor eccentricity where the rotor is displaced vertically toward the U2 coil by 13 % of the normal air-gap length. In addition, the experimental IPM has a maximum erection tolerance of 7 % and hence the total eccentricity value ranges from 13 up to 20 %. III. VOLTAGE DIFFERENCE UNDER ECCENTRICITY This paper introduces a theoretical expression of the circulating current for the purpose of comparing connected and unconnected neutral points. The circuit diagram of parallel connected stator winding based on the one reported in [8] is shown Fig", + " The following considers the induced voltage at each stator winding under static eccentricity. Fig. 5 shows simplified diagrams of the motor under eccentricity. When the rotor is under static eccentricity, the air-gap of each tooth can be expressed as \u03b4 (\u03b8t) = \u03b4gap - (\u03b1cos(\u03b8t)+ \u03b2sin(\u03b8t)), (3) where \u03b4gap is the average length of the air-gap in the case of no eccentricity and \u03b4 is the relative value of static eccentricity. Then, \u03b8t is the mechanical angle of the tooth and (\u03b1, \u03b2) is eccentricity value in the coordinate system. In this case, according to Fig. 3, each tooth is placed every 60 degrees. Therefore, each air-gap, i.e., \u03b4v1, \u03b4v2, \u03b4w1, and \u03b4w2, are defined as \u03b4v1 = \u03b4w2 and \u03b4v2 = \u03b4w1. Next, we calculate the flux density difference caused by static eccentricity because the induced voltage difference is proportional to the flux density difference. In this case, it is assumed that iron permeability is infinite and that eddy current does not occur. The flux density in the air-gap B\u03b4 at the no-load external drive is written as PMPMba PMr h hBB 2)( 2 ++ = \u03bc\u03b4\u03b4\u03b4 , (4a) where hPM is the height of the magnet, Br is the remanent flux density, and \u03bcPM is the recoil permeability" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002717_0959651812443925-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002717_0959651812443925-Figure5-1.png", + "caption": "Figure 5. The pole placement region.", + "texts": [ + "4,10 According to the approach outlined above in subsections \u2018\u2018Affine parameter-dependent system\u2019\u2019 and \u2018\u2018PMSM T\u2013S fuzzy model\u2019\u2019, the local linear model matrices for the nonlinear augmented states of the IPMSM at the ith selected operating point are obtained as follows A02 = 0 1 0 0 0 0 0 0 0 1860 0 0 0 0 0 0 0 0 0 20 82119 0 0 0 1 0 0 0 0 0 0 0 1 0 0 0 2 6666664 3 7777775 A03 = 0 1 0 0 0 0 0 0 1145 1860 0 0 0 49 0 0 0 0 0 20 0 0 0 0 1 0 0 0 0 0 0 0 1 0 0 0 2 6666664 3 7777775 A04 = 0 1 0 0 0 0 0 0 0 1860 0 0 0 0 0 0 0 0 0 20 0 0 0 0 1 0 0 0 0 0 0 0 1 0 0 0 2 6666664 3 7777775 C= 0 0 0 0 1 0 0 0 0 0 0 1 1 0 0 0 0 0 0 0 1 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 2 6666664 3 7777775 D= 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 1 0 0 0 0 0 1 2 6666664 3 7777775 Next, in order to design state-space feedback gains Ki for each subsystem, the following steps are undertaken. 1. Specify the LMI region, equation (6), in order to place the closed-loop poles in this region (pole placement) and also to guarantee some minimum decay rate and closed-loop damping. The mentioned region is shown in Figure 5, as the intersection of the half-plane x \\ 25 and of the sector centered at the origin and with inner angle 2p/3. 2. Choose a four-entry vector specifying the H2/HN cost function, equation (1): [g0 n0 a b]=[0 0 1 1]. 3. Minimize the H2/HN cost function, equation (1), subject to the mentioned pole placement constraint by using equations (4), (5), (8) and (9). The local H2/HN control gain matrices Ki, the final fuzzy H2/HN controller matrix Kfuzzy and the overall fuzzy model matrix Afuzzy are obtained as follows K1 =104 0:0033 0:0011 0 0 0:0097 0:0002 0:0002 0:0001 0:0386 0:0009 0:1467 3:8313 2 6666664 3 7777775 T Parameter Value Rs 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.15-1.png", + "caption": "FIGURE 6.15 A plain surface, sliding, or journal bearing.", + "texts": [ + " The two surfaces slide over each other, and this motion can be facilitated by means of a lubricant that gets squeezed by the motion of the components and can under certain conditions generate sufficient pressure to separate the surfaces, thereby reducing frictional contact and wear. A typical application of sliding bearings is to allow rotation of a load-carrying shaft. The portion of the shaft at the bearing is referred to as the journal, and the stationary part, which supports the load, is called the bearing (see Figure 6.15). For this reason, sliding bearings are 204 Rotating Flow often collectively referred to as journal bearings, although this term ignores the existence of sliding bearings that support linear translation of components. Another common term is plain surface bearings. This section is concerned principally with bearings for rotary motion, and the terms journal and sliding bearing are used interchangeably. There are three principal regimes of lubrication for sliding bearings: 1. boundary lubrication 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003334_978-3-7091-1379-0_43-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003334_978-3-7091-1379-0_43-Figure3-1.png", + "caption": "Figure 3. Left: results of the skin spatial calibration process. Right: the actual body part to be calibrated.", + "texts": [ + " Skin SW technologies include algorithms and data structures allowing tactile data to travel from the lowest level (i.e., the actual sensors) up to user level applications. Both general-purpose and robot-specific tactile data processing architectures have been investigated and experimentally evaluated in terms of real-time performance, specifically taking into account bandwidth, jitter and reliability issues. Specific emphasis has been put on the so-called skin spatial calibration problem, i.e., the problem of self-estimating the location of tactile elements mounted on a robot body part, see Figure 3. To this aim, different solutions have been investigated (Cannata et al., 2010a; Prete et al., 2011; McGregor et al., 2011). Result 4. Implementation of an architecture for touch-based social interaction, and development of classification algorithms for touch-based social interaction. Robot skin makes the design and implementation of a number of robot behaviours possible both at the reflexive and purposive levels. On the one hand, skin based protective reflexes have been developed, which are based on reflex receptive fields as reported in studies from human subjects (Pierris and Dahl, 2010)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.11-1.png", + "caption": "Fig. 2.11 The model of the Graham escapement", + "texts": [ + " Besides the Graham escapement, he was also the inventor of the mercury compensation pendulum, the Cylinder escapement for watches and the first chronograph. The mercury pendulum can achieve an accuracy of within a few seconds per day, a monumental achievement for the time. Graham refused to patent these inventions because he felt that they should be used by other watchmakers as well. He was a truly talented and generous inventor (Fig. 2.10). The Graham escapement is also called the Deadbeat escapement. It is a modified version of the Anchor escapement and mostly eliminates the aforementioned coil problem. Figure 2.11 shows the model of the Graham escapement. Similar to the Anchor escapement, it mainly consists of an escape wheel, a pallet fork and a pendulum. During the operation, the escape wheel is driven by the power train and moves clockwise. On the other hand, the pallet fork and the pendulum are joined together and swung. Graham made a number of delicate modifications. First, the anchor pallet is concentric to its center. Second, the tip of each limb of the pallet has a specific shape designed to provide an impulse as the escape wheel tooth slides across the surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002680_10402004.2012.664836-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002680_10402004.2012.664836-Figure2-1.png", + "caption": "Fig. 2\u2014Photograph of the bearing test rig. (color figure available online.)", + "texts": [ + " [1] and [2], and for the case of identical materials (steel 1 and steel 2) either such of the interface (z = z1 = z2), the film thickness becomes h = 2\u03c1c2 \u03c9z \u221a |R|2 1 \u2212 |R|2 [3] Equation [3] is also known as a quasistatic spring model (Drinkwater, et al. (7)). Experimentally, the reflection coefficient is determined by dividing the reflected amplitude for an oil film (i.e., R = Rfilm in Fig. 1b) by the reflection amplitude for an air interface (i.e., R = Rair in Fig. 1a). This is because with a steel\u2013air interface virtually the entire ultrasonic wave amplitude is reflected and so the incident signal is equivalent to Rair. Hence, R(f ) = Rfilm(f ) Rair(f ) [4] Test Apparatus Figure 2 shows the rolling bearing test rig used in this work. The bearing is rotated by an induction motor with a speed controller through a pulley and belt system. The maximum speed for speed controller is 1,500 rpm. The lubricating oil is recirculated at a flow rate of 3.5 mL.s\u22121 using a peristaltic pump. The load was applied using a hydraulic cylinder that can provide a maximum load up to 200 kN. The hydraulic load acts as a lever arm such that the maximum loaded point is at the top of the bearing. Figure 3 shows a schematic of the bearing and the location of the transducer" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003251_s1000-9361(11)60092-7-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003251_s1000-9361(11)60092-7-Figure1-1.png", + "caption": "Fig. 1 Missile pitch plane dynamics.", + "texts": [ + " The proposed longitudinal controller offers the greatest scope for controller improvement, and guarantees damping ratio \u03b6 > 0.7 with overshoot <10% in the presence of dynamic uncertainty and large disturbance. This approach is tested through various scenarios for the validated nonlinear dynamic flight model of the real ballistic missile system with autopilot exposed to external disturbances. This section demonstrates the longitudinal dynamics of the existing ballistic missile system during the boost trajectory. Fig. 1 shows the missile pitch plane dynamics. Where \u03b1 is the angle of attack, (\u00b0); m the total missile mass, kg; VM the missile total velocity, m/s; \u03b8 and \u03d1 are the flight path angle and missile pitch angle respectively, (\u00b0). For the system under investigation, the missile has four air rudders arranged, as shown in Fig. 2, where \u03b4i (i=1,2,3,4) is rudder deflection angle. For a non-perturbed ascent trajectory (\u201cboost tra- jectory\u201d), there is a trajectory that results from standard predicted values of missile thrust, weight, lift, drag, and that experiences no wind velocity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002790_s11465-013-0254-x-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002790_s11465-013-0254-x-Figure4-1.png", + "caption": "Fig. 4 Hertzian deformations on the cam (1) and the roller (2)", + "texts": [ + " This distance is defined by the following equation [11,12]: D\u00f0t\u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi L21 \u00fe L3 \u00fe LC\u00f0t\u00de 2 \u2013 2L1 L3 \u00fe LC\u00f0t\u00de cos\u03b2\u00f0t\u00de r , (9) with LC\u00f0t\u00de \u00bc L3$ \u03c93\u00f0t\u00de \u03c91 \u2013\u03c93\u00f0t\u00de , (10) \u03c93(t) is the time angular velocity of the oscillating roller follower (3). 3.2 Expression of the first hertzian stiffness Kh1 The contact area between the roller (2) and cam (1) is characterized by an elastic non linear hertzian stiffness which depends on the position of the instantaneous contact point C1. The general form of this stiffness is as follows [13]: Kh1 \u00bc Fc1 hc\u00f01\u00de \u00fe hr\u00f02\u00de , (11) hc and hr characterize the depth of hertzian deformations on the cam (1) and the roller (2) respectively (Fig. 4). The localized deformation hr and hc are defined by: hr\u00f02\u00de \u00bc Rr\u00f02\u00de \u2013 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R2 r\u00f02\u00de \u2013 g 2 1 q , (12) hc\u00f01\u00de \u00bc Rc\u00f01\u00de \u2013 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R2 c\u00f01\u00de \u2013 g 2 1 q , (13) Rr(2) is the constant radius of the roller (2). Rc(1) is the radius of curvature of the cam profile which varies with the position of the contact point and is given by the following expression [13]: Rc\u00f01\u00de \u00bc Rb\u00f01\u00de \u00fe Rr\u00f02\u00de \u00fe H\u00f0 \u00de 2 \u00fe _H \u00f0 \u00de \u03c91 2 3 2 Rb\u00f01\u00de \u00fe Rr\u00f02\u00de \u00fe H\u00f0 \u00de 2 \u00fe 2 _H \u00f0 \u00de \u03c91 2 \u2013 Rb\u00f01\u00de \u00fe Rr\u00f02\u00de \u00fe H\u00f0 \u00de \u20acH \u00f0 \u00de \u03c92 1 \u2013Rr\u00f02\u00de, (14) Rb(1) is the base circle radius of the cam (1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001969_1.3605512-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001969_1.3605512-Figure1-1.png", + "caption": "FIG. 1. a Schematic of hollow microfiber generation using a two-channel microfluidic device. b Photograph of the microfluidic device showing generation of alginate microfibers composed of 1% w/v Na-alginate mixed with blue dye. c Enzyme-entrapped alginate hollow fibers.", + "texts": [ + " VBPO The VBPO assay was based on the method of de Boer et al.,29 and involves bromination of Phenol Red to Bromophenol Blue by monitoring the conversion at 580 nm. In brief, each assay contained 2 ml of 100 M phenol red, 0.2 ml of 1.0 M potassium bromide, 0.2 ml of 10 mM hydrogen peroxide and enzyme 0.1 ml . At intervals of 1 min, absorbance was measured using an AvaSpec-3648 Fiber Optic Spectrometer Eerbeek, The Netherlands , scanning the absorbance between 350 and 650 nm. The enzyme concentration was 4.5 units/ml. Figure 1 a is a schematic of hollow microfiber generation and also shows the microfluidic device used. Fabrication of the microfluidic device was carried out as described in our previous reports.30,31 In brief, a PDMS platform with a preformed principal hole 1 mm inner diameter was employed and pulled borosilicate glass micropipettes were used. A photograph of the working microfluidic device containing hollow microfibers is shown in Fig. 1 b . For the fabrication of alginate hollow microfibers, two fluids 2% w/v sodium alginate for sheath flow and 100 mM CaCl2 for core flow were introduced into the sheath and core inlet ports, respectively, using two syringe pumps. A three-dimensional 3D coaxial sample flow stream of alginate, around the core flow of CaCl2 solution, was formed at the point where the two flows merged. At the fluid interface, the sodium alginate solution met the polycationic Ca2+ material and gelation of cylindrical hollow fibers was achieved by diffusion-controlled ionic cross-linking", + " Finally, the fiber was embedded into half-cured PDMS solution and subsequent curing was carried out on a hot plate for 1 h at 70 \u00b0C. The reaction mixture was composed of 0.025 mM Phenol Red, 0.5 mM potassium bromide, and 0.25 mM hydrogen peroxide in 50 mM MES buffer, pH 6.5. Reactants were thoroughly mixed and 0.1 ml of enzyme solution was added to start the reaction. Reactions were monitored at 1 min intervals, with subsequent scanning of the absorbance at 350\u2013650 nm. With continued Bromophenol Blue production, the peak at 590 nm increased continuously until the reaction was complete at 170 min Fig. 2 . As shown in Fig. 1 b , the successful generation of hollow alginate microfibers is performed by the coaxial microfluidic device Fig. 1 a . The alginate hollow fibers were highlighted using a blue dye Papicel Red IJ-F3B; Eastwell Co. Ltd., Seoul, Korea that stains the alginate. In initial experiments, hollow alginate microfibers were formed without enzyme. At a fixed sheath flow condition, based on the increase of the core flow rates, the diameters of the hollow fiber were increased by approximately 45%. Also, to demonstrate a symmetric hollow-fiber shape, fluorescein isothiocyanate-bovine serum albumin FITC-BSA was loaded and immobilized into the outer-wall of the alginate hollow fibers and imaged using confocal laser scanning microscopy LSM-510 META instrument; Carl Zeiss, Darmstadt, Germany , scanning electron microscopy This article is copyrighted as indicated in the article" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000165_bf00935252-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000165_bf00935252-Figure8-1.png", + "caption": "Fig. 8. CCS for a more general payoff.", + "texts": [ + " Since the payoff, dynamics, and constraints are the same as in Section 2, the Hamiltonian (15), multiplier conditions (16), and adjoint equations carry over. The speed adjoints still vanish at t r. When the termination is on the circular arc AB, the complete analysis of Section 2 is valid, and the ETMs can be drawn for both players subject to the proviso that tBlrt<~6. (49) For termination on the sides PA and PB of the fan, all adjoints except the pursuer's angular adjoint and the Hamiltonian vanish at termination. The CCS is chosen as shown in Fig. 8 with Of = \u00b190 \u00b0 for PA and PB, respectively. The equations for the game split into two independent sets for the sides of the fan also. For each pair of rr, M~f values, the pursuer's ETM consists of two trajectories with r i l i= 90\u00b0- 3, o)~r<0, and ri~\u00a2 = -90\u00b0+~ , w l f > O , respectively. Variation of r r over [0, R] affects the switching of ~'i and aJi. This, together with the portion of the ETM where friazl<6, gives the pursuer's total ETM. The evader's ETM remains unchanged for the sides of the fan", + " Returning to the circular capture set of Section 2, a more general payoff J = ~bl(01r) + ~b2(4'2r) + dt (50) is considered. A similar payoff is used in Ref. 15 for the game of two cars. Since the integral portion of the payoff, the dynamics, and constraints are unchanged from those of Section 2, the Hamiltonian, multiplier conditions, and adjoint equations carry over. The Hamiltonian and speed adjoints vanish at t~. The angular adjoints are not necessarily zero at termination. The CCS is chosen as in Fig. 8. Here, 0y is a function of both the terminal angular adjoints. Although the solution in the small cannot be described in terms of the individual players' ETMs, simplicity of derivation and computational economy results. For planar pursuit--evasion with variable speeds, a particular real-space representation of the game is open-loop optimal. The solution in the small is discussed in terms of the players' ETMs. The ETM contains all the maneuvers that may be executed by a player in any encounter. It is the same whether he is pursuing or evading" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001367_bf01231426-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001367_bf01231426-Figure1-1.png", + "caption": "Fig. 1. Spring-pendulum configuration.", + "texts": [ + " Another feature of the spring-pendulum equations of motions is, that they are not complicated by rotating axes and the corresponding Coriolis forces. With all those advantages in mind it was an easy task for us to study numerically the manifold of symmetric periodic solutions of this problem. We found that there is a very large variety of periodic resonance solutions. We will use a (righthanded) coordinate system (x, y) which is such that the x-axis points downwards, in the direction of the constant acceleration g of gravity (Figure 1). We will also assume that the natural length (when no traction or force is applied) of the spring is lo > 0 and also that the spring is purely linear, with spring constant k > 0. In other words if r is the radius vector (distance of the pendulum bob from the origin or support), the stretching of the spring is r-lo and the restoring spring force is - k ( r - 10). The potential energy is thus composed of the elastic energy k ( r - lo)a/2 and the gravitational energy - m g x , so that the Lagrangian of this problem (per unit mass) is: ~o = T + U = 89 (22 + " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001236_1.4001281-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001236_1.4001281-Figure1-1.png", + "caption": "Fig. 1 Three dynamic walking models capable of show with hip and knee \u201eif applicable\u2026 actuation to recover fro the swing and stance leg. Knee actuation produces a t simplest walking model by Garcia et al. \u202024\u2021. All variab length L, hip mass M, and gravitational constant g as b with the hip mass. \u201eb\u2026 The straight-legged equivalent o McGeer \u202029\u2021. Both models have anthropomorphically di the torso. See Table 1 for the model parameter descrip", + "texts": [ + "1 The Walking Models. A total of three walking models ere used throughout this simulation study. All models showed a yclic passive dynamic gait when powered only by gravity using a light downhill slope. We extended these passive walking models ith actuation in the hip joint and knee joint if present , but this ddition was only used when a recovery action was required in esponse to a stumble. To reveal the very fundamental principles of stumble recovery, e primarily used the \u201csimplest walking model\u201d 24 Fig. 1 a . his irreducibly simple walking model was based purely on mehanical principles. Despite its simple nature, the dynamic priniples of this model have proven to provide useful insight into rinciples of human gait 25\u201327 . Furthermore, similar simple odels were found to adequately describe the dynamics of the uman body during balance recovery 15,28,14 . It consisted of wo rigid links, connected by a frictionless rotational joint at the ip. Point masses were located at the hip and at the feet. The asses at the feet were infinitesimally small compared with the ass at the hip, which resulted in a dynamic motion of the hip that s not influenced by the swing leg dynamics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000654_s11071-009-9482-3-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000654_s11071-009-9482-3-Figure6-1.png", + "caption": "Fig. 6 Rotor-bearing model with radial clearance", + "texts": [ + "1 Rotor model and equations of motion In the experimental setup, the lower bearing can move in the clearance and the lateral oscillation and the inclination oscillation of the rotor couple each other. This rotor system can be approximated by a model with 8 degrees of freedom. Since the analysis of such a system is laborious and the only phenomena related to the lateral modes corresponding to those with frequencies p2 and p3 in Fig. 2 are observed in experiments, we adopt a symmetrical system shown in Fig. 6 where the rotor is mounted at the center of the shaft. If we consider only symmetrical lateral oscillations, that is, the case that the deflections of the upper and lower bearings are the same, it has two natural frequencies pf > 0 and pb < 0. However, since the relationship pf = \u2212pb called internal resonance holds and several nonlinear resonances occur in the same rotational speed range in this case, the nonlinear resonance phenomena may becomes more complicated ones and different from the case of the experimental setup" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002941_978-3-642-36279-8_13-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002941_978-3-642-36279-8_13-Figure3-1.png", + "caption": "Fig. 3 This figure illustrates the arrangements of nodes and beams attached to a horizontal beam \u03c4\u2217, which is being currently placed. The arrangement of nodes and beams for a vertical beam \u03c4\u2217 can be found by rotating the the figure by 90 degrees clockwise. This figure is used in conjunction with the rules of the Next Beam algorithm.", + "texts": [ + " The vertical columns do not cause deadlocks and can be placed in any order. Hence, we focus on the construction of each stratum. The DATS algorithm approaches the stratum construction problem on a beam-by-beam basis. During each iteration, the DATS algorithm examines the local state of each available beam. Available beams are those unbuilt beams in the current stratum that are immediately adjacent to the currently built structure. We define the adjacent beams \u03c4i and nodes Ni of \u03c4\u2217 as the six beams and nodes that are positioned relative to \u03c4\u2217 as shown in Fig. 3. Thus at every iteration, the DATS algorithm has a current list of available beams. Using the rules of the Next Beam algorithm (Alg. 2), the DATS algorithm determines whether the beam \u03c4\u2217 can be placed at all and if it should be placed as a beam \u03c4\u2217 or module \u03c4\u2217Ni. In the algorithm, A, B, and C refer to the columns and rows illustrated in Fig. 3. We denote that a beam or node is occupied O(\u2217) or empty E(\u2217) where \u2217 can either be \u03c4i or Ni. Further, O(\u2217) also denotes that a row or column contains a node when \u2217 is A, B, or C and similarly for E(\u2217). In Alg. 2, we use the notation &, |, \u2295, and (\u2217) to represent the boolean operations AND , OR, XOR, and NOT , respectively. After a beam has been placed, a list of available beams is generated and sorted in any manner. The manner in which beams are sorted has no bearing on the correctness of the DATS algorithm and only serves to increase performance as discussed in Sec", + " The only exception to this condition is Rules 5 of the Next Beam algorithms because the placement of only one beam can never result in an additional node. Algorithm 1. Distributed Assembly for Truss Structures (DATS) Algorithm (BEAMS is a list of unbuilt beams, which are immediately adjacent to the currently built stratum) 1: BEAMS\u2190 A seed beam in the current stratum 2: while not finished do 3: Next Beam Algorithm\u2190 \u03c4\u2217 4: All beams immediately adjacent to previously placed beams\u2192 BEAMS 5: Sort BEAMS Algorithm 2. Next Beam Algorithm (A, B, and C denote the row and columns shown in Fig. 3) Require: \u03c4\u2217 1: if O(N2)&(E(B)|(O(\u03c41)\u2295O(\u03c45))) then 2: Place module \u03c4\u2217N1 {Rule 1} 3: else if O(N1)&(E(C)|(O(\u03c42)\u2295O(\u03c44))) then 4: Place module \u03c4\u2217N2 {Rule 2} 5: else if ((E(C)|(O(\u03c42)\u2295O(\u03c44)))&(O(\u03c43)|E(A))) then 6: Place module \u03c4\u2217N2 {Rule 3} 7: else if ((E(B)|(O(\u03c41)\u2295O(\u03c45)))&(O(\u03c46)|E(A))) then 8: Place module \u03c4\u2217N1 {Rule 4} 9: else if O(N2)\u2295O(N1) then 10: Place module \u03c4\u2217 {Rule 5} 11: else if (E(\u03c41)&E(\u03c42)&E(\u03c43)&E(\u03c44)&E(\u03c45)&E(\u03c46)) then 12: Place module \u03c4\u2217N2 {Rule 6} We complete the discussion of the DATS algorithm by stating some properties and theoretical guarantees", + " After the first part is placed, all other parts are attached to previously placed parts. Theorem 1. The DATS algorithm can correctly place parts without causing any inconsistencies. Proof. We must show that the Next Beam algorithm (Alg. 2) does not cause inconsistencies i.e. deadlocks both locally and globally. We first consider the local implications of the Next Beam algorithm. In order to ensure that there are no unplaced parts between previously placed parts, we must verify that the following conditions are satisfied based on Fig. 3. Pictorially, we illustrate the six situations, which violate these conditions, in Fig. 4. (O(N1)|O(\u03c46))&(O(N2)|O(\u03c43))&E(\u03c4\u2217) (1) E(N1)&O(\u03c41)&O(\u03c45), E(N2)&O(\u03c42)&O(\u03c44) These conditions are trivially satisfied at the first iteration (k = 0) because we start with a single node. We must show that if iteration k satisfies these conditions, by building \u03c4\u2217, \u03c4\u2217N1, or \u03c4\u2217N2. We consider the two cases in which a horizontal \u03c4\u2217 shown in Fig. 3 becomes either \u03c46 for a subsequent horizontal \u03c4 \u2032\u2217 or \u03c44 for a subsequent vertical \u03c4 \u2032\u2217 in the vertical configuration of. It can be shown exhaustively that any action (the placement of \u03c4\u2217, \u03c4\u2217N1, or \u03c4\u2217N2) dictated by the Next Beam algorithm will result in the satisfaction of the conditions. Thus by mathematical induction, we can show that the Next Beam algorithm does not cause local inconsistencies. Similarly, we define a global inconsistency as the independent placement of a part in a continguous row or column, which already contains a part", + " Unless a 2-D stratum is finished, there is at least one part placement that can occur. Proof. The proof follows from Thm. 1. For a partially built structure, there is at least one unplaced beam or module. When there is exactly one unplaced part, this part can be placed using Rules 1 \u2212 2 of the Next Beam algorithm. We must show that this part can indeed be placed by these rules. Since all prior parts have been placed using the rules of the Next Beam algorithm and these rules cover the span of all possible valid configurations of beams and nodes illustrated in Fig. 3, it must be true that any configuration of the single unplaced part can be placed by the Next Beam algorithm. When there are two or more unplaced parts, we must show that at least one of these parts must be in a configuration, which can be completed with the Next Beam algorithm. For these cases, the unplaced beams/modules can be categorized into three groups: parts, which can be placed immediately, parts, which can be placed but not at the current moment because the placement would eventually result in a deadlock, and parts, which cannot be placed due to an inconsistency" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000174_iros.2008.4650826-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000174_iros.2008.4650826-Figure2-1.png", + "caption": "Fig. 2. (a) The humanoid robot ARMAR-III moving in a translational dimension. (b) The effect in workspace when changing the C-space value for the dimension associated with the torso pitch joint.", + "texts": [ + " Therefore, an upper bound for the workspace movements of each limb is used for an efficient and approximated uniform sampling. A change \u03b5trans in a translational component of the Cspace moves the robot in workspace by \u03b5trans millimeters. All other dimensions of the C-space have to be investigated explicitly to derivate the upper bound of the robot\u2019s workspace movement. Table I gives an overview of the maximum displacement of a point on the robot\u2019s surface when changing one unit in C. The effects of moving one unit in the different dimensions can be seen in Fig. 2. The different workspace effects are considered by using a weighting vector w whose elements are given by the values of the workspace movements from table I. In Eq. 1 the maximum workspace movement dWS(~c) of a C-space path ~c = (c0, ...,cn\u22121) is calculated. dWS(~c) = n\u22121 \u2211 i=0 wici (1) To sample a C-space path between two configurations ~c1 and ~c2, the vector ~vstep is calculated (Eq. 2). For a C-space displacement of ~vstep it is guaranteed that the maximum workspace displacement is 1mm. ~vstep(~c1,~c2) = (~c2 \u2212~c1) dWS(~c2 \u2212~c1) (2) The maximal workspace stepsize \u03b5ws can be specified in millimeters, which allows to control the granuality of the planning algorithms in an easy way" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001491_09544062jmes1452-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001491_09544062jmes1452-Figure4-1.png", + "caption": "Fig. 4 Operating principle", + "texts": [ + " For reaching the effective inlet and outlet ports during operation, the design of ports JMES1452 \u00a9 IMechE 2009 Proc. IMechE Vol. 223 Part C: J. Mechanical Engineering Science at BRIGHAM YOUNG UNIV on June 25, 2015pic.sagepub.comDownloaded from becomes important. First, the distribution angle of the inlet port should be bigger than the outlet port, which can make an internal pressure and increase the air-extracting ratio. If the design is shown in Fig. 3, the operation principle of intake and exhaust is shown in Fig. 4. First, the two rotors rotate in opposite directions. Figures 4(a) to (f) represent a cycle rotating. When the rotors rotate as given in Fig. 4(a), the inlet and outlet ports are sealed. Coming to Figs 4(b) and (c), the angle range of the inlet port and the capacity would increase. This is the extracting process and it is a low-pressure area. At the outlet port, the capacity may change from large to small, so the air would be pressed. At the time, the outlet port is still sealed. When continuing rotating to Fig. 4(d), the angle of the inlet port is becoming small until the whole intake operation is finished. At this time, the outlet port is not sealed anymore. Then it begins to exhaust while the capacity becomes small, the outlet port is a high-pressure area. The inlet and outlet ports would be sealed again till Fig. 4(e). In For a clearer concept of the design, Fig. 5 shows the rotor motion locus. Because the parameter \u00b5 is set 0.8, there is no mesh in the crown and the root of the rotor profile. If the ports of the inlet and outlet are not sealed at these positions (see Figs 4(e), (f), and (a)), the gas may leak out through these two ports. Hence, this would affect the pump performance. If the inlet and outlet design is as shown in Fig. 3, some high-pressure air is represented as the dark area in Fig. 4(e) may go through Figs 4(e) and (f) to Fig. 4(a), finally it would be brought to the inlet area. Under a fixed value \u03b4, \u00b5 is taken as 0.7, 0.8, 0.9, and 1. Here, \u03b4 is given as 1.5, 1.6, and 1.7. If r = 40, the rotor profile can be derived from the above equation. Figure 6 shows results for different \u00b5 (\u03b4 = 1.5). It is obvious from the figure that the distance is becoming far from one claw tip of the rotor to the other rotor profile when \u00b5 becomes small. However, the inlet and outlet ports design is given in Fig. 3. If the angle of the Proc. IMechE Vol. 223 Part C: J. Mechanical Engineering Science JMES1452 \u00a9 IMechE 2009 at BRIGHAM YOUNG UNIV on June 25, 2015pic.sagepub.comDownloaded from port is too large, then some air may be produced from the outlet port to the inlet port (see Fig. 4(e)). This may lead to a low air-extracting efficiency. If the port is like the design of Fig. 3, the carryover area can be defined as Ac, showed as Fig. 7. Besides, for evaluating the area efficiency of the rotor, this article may call it as an invalid area, which means when the inlet and outlet ports are sealed, it is useless for extracting the air no matter how much area is present between the two rotors. But, for a convenience to compare different parameters, this aricle would take the invalid area Anf as three different positions (see Fig", + " The clearance may also affect the pump efficiency. To prevent air leaking from the clearance, usually at the top of the rotor an arc whose Fig. 13 Design of the outlet port (\u00b5 = 1) radius is close to the chamber radius would be used. The wrap angle of the arc depends on the working condition. For a better pump performance, the inlet and outlet ports design are also important. For example, if the parameter is set as \u00b5 = 1 and the positions of the inlet and outlet ports are same as Fig. 3. When two rotors rotate to the position like Fig. 4(d), the inlet port is closed and the outlet port is open (high-pressure area) to exhaust gas until the position of Fig. 4(e). Some highpressure gas would be brought from Figs 4(e) to (f). However, at the position of Fig. 4(f) (same as Fig. 6, \u00b5 = 1), the high-pressure area would be divided into three sections that may produce large noise, and also make invalid work. To solve the problem, the size and the position of the outlet port can be designed as Fig. 13 for reducing the gas pressure. Thus, lowering the pressure of the outlet port may improve the noise and wasted work problem. Though \u00b5 = 1 can make bigger area efficiency, it has some shortcomings. Hence, the value \u00b5 is suggested to be set <1. This article presents a simple parameter design of the claw-type rotor profile. Different parameter designs would affect the gas port design. If the design of \u00b5 is <1, the design of the gas port should avoid the high-pressure gas going through the clearance from the outlet port to the inlet port between the two rotors. Therefore, when two rotors rotate to the position of the intersection point of the chamber profile (like Fig. 4(a) or Fig. 4(e)), the design of the inlet and outlet ports should be sealed. However, when \u00b5 equals 1, in order to lower the noise and wasted work, the size of the gas ports and their positions need to be adjusted. In addition, this article also derives the feasible design range of \u00b5 and \u03b4, which make no interference in the profile of the two rotors. Lastly, the parameter design shows that area efficiency would increase when the value of Proc. IMechE Vol. 223 Part C: J. Mechanical Engineering Science JMES1452 \u00a9 IMechE 2009 at BRIGHAM YOUNG UNIV on June 25, 2015pic" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001368_j.mechatronics.2010.09.010-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001368_j.mechatronics.2010.09.010-Figure5-1.png", + "caption": "Fig. 5. A desc", + "texts": [ + " It may happen that the descent stop point occurs later (at a greater distance) than the ascent start point as shown in and therefore H1 \u00bc b tan h1 tan h2 tan h1 \u00fe tan h2 : H2 is calculated as: H2 \u00bc g tan h1 tan h2 tan h1 \u00fe tan h2 : \u00f08\u00de New node points \u00f0dnew 2 ; hnew 2 \u00de and \u00f0dnew 3 ; hnew 3 \u00de can be found as: dnew 2 \u00bc d2 H1 \u00fe H2 tan h2 ; hnew 2 \u00bc h2 \u00fe \u00f0H1 \u00fe H2\u00de; dnew 3 \u00bc d3 \u00fe H1 \u00fe H2 tan h1 ; hnew 3 \u00bc h3 \u00fe \u00f0H1 \u00fe H2\u00de; \u00f09\u00de Old node points (d2,h2) and (d3,h3) are replaced with new node points \u00f0dnew 2 ; hnew 2 \u00de and \u00f0dnew 3 ; hnew 3 \u00de in the node list. A different scenario occurs when a descent\u2013ascent gap exists but is less than the specified value g. Now we raise the descent stop and ascent start points along their slope lines as follows, see Fig. 5. Here b \u00bc d3 d2; c \u00bc H tan h2 ; e \u00bc H tan h1 and it is seen from Fig. 5 that: g \u00bc b\u00fe c \u00fe e \u00bc \u00f0d3 d2\u00de \u00fe H tan h2 \u00fe H tan h1 : \u00f010\u00de Hence H can be found as: H \u00bc \u00f0g d3 \u00fe d2\u00de tan h1 tan h2 tan h1 \u00fe tan h2 ; \u00f011\u00de and the new node points are: dnew 2 \u00bc d2 H tan h2 ; hnew 2 \u00bc h2 \u00fe H; dnew 3 \u00bc d3 \u00fe H tan h1 ; hnew 3 \u00bc h3 \u00fe H: The new node points now replace the old ones in the node list. Note that such scenarios cannot arise in an \u2018ascent-followed-by-descent\u2019 case because of the way the ascent and descent nodes are defined. 3500 Stair algorithm (2.5 km segment length) Terrain Altitude Computed Trajectory Step 6: For the last node point (say F) which could be the landing point, find the intersection of a line (with the vehicle\u2019s descent rate as its slope) joining F to the previous section, and if an intersection with the previous section does not exist, move to the section before it and so on until an intersection is found" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure3-1.png", + "caption": "Fig. 3. Angular misalignment in the cylindrical, revolute and translation joints.", + "texts": [ + " (4) as P P j j j j i i i i= + \u2212 \u2212d r A \u03c9 s r A \u03c9 s& & % & % (6) where \u03c9 is the angular velocity of body i. When the bushing joint is not located in the center of mass of the connected bodies the transport moments are ( ) ( ) ' ' T P i i i i T P j j j j = = n A s f n A s f % % (7) where P is% is a skew-symmetric matrix made with the components of vector P is . The degrees of freedom to be penalized by this joint are the normal translational displacement, dn, depicted in Fig. 2, and the angular displacement due to the misalignment of the vectors, presented in Fig. 3. Based on Fig. 2 the misalignment of the axis of the joint in the tangential direction is given as ( )( )T T t j j j=d d s A s (8) while the misalignment normal to the joint axis in body j is n t= \u2212d d d (9) The unit vector si defines the joint axis in body i and the unit vector sj defines the joint axis in body j. The forces due to the normal misalignment of the axis are 1[ ( ) ( ) ]i n n n n n n j i K f b d \u03b4 \u03b4 \u03b4 \u03b4= \u2206 + + = \u2212 f d f f & (10) In Eq. (10) the magnitude of the deformation T n n n\u03b4 = d d and the velocities n\u03b4& and t\u03b4& are T n n n n\u03b4 \u03b4= d d& & (11) T t t t t\u03b4 \u03b4= d d& & (12) n t= \u2212d d d& & & (13) ( )( )T T t j j j=d d s A s& & (14) The moment due to the angular misalignment of vectors si and sj requires calculating angle \u03b8 between them" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003460_1350650112466768-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003460_1350650112466768-Figure1-1.png", + "caption": "Figure 1. Schematic drawing of the piston\u2013liner system.", + "texts": [ + " So, a comprehensive lubrication model should be built to evaluate the influence of various deformations of the piston skirt and the liner on the piston\u2013liner system lubrication. As an effort to reach this, it will be discussed in details that the calculation of the deformations and the influences of the deformations on the piston secondary motion, the oil film pressure and thickness, and the friction loss of the piston skirt\u2013 liner system in this study. A schematic drawing of the piston\u2013liner system is given in Figure 1, and all the forces and moments acting on the piston are shown in Figure 2. The equations of motion are derived based on the dynamic equilibrium of all the forces and moments applied to the piston. The system equation can be expressed as the following mpin\u00f01 a L\u00de \u00fempis\u00f01 b L\u00de mpin a L\u00fempis b L Ipin pis L \u00fempis\u00f0a b\u00de\u00f01 b L\u00de mpis\u00f0a b\u00de bL I pin pis L 2 4 3 5 \u20acet \u20aceb \" # \u00bc F \u00fe Fs M\u00feMs \" # \u00f01\u00de where Fs \u00bc X3 i\u00bc1 Ffi \u00f0Fg \u00fe Fip \u00fe Fic \u00fe Fr\u00feFf \u00de sin \u00fe fp sgn\u00f0 _ \u00de cos cos fp sgn\u00f0 _ \u00de sin \u00f02\u00de Ms \u00bcMf\u00feFgCp FicCg\u00fe X3 i\u00bc1 FfiLi \u00feFrCp \u00f0Fg\u00fe Fic\u00fe Fip\u00feFr\u00feFf \u00de fp sgn\u00f0 _ \u00de rpin cos fp sgn\u00f0 _ \u00de sin \u00f03\u00de fp is the friction coefficient of the wrist-pin bearing. Ffp is assumed as a linear function of the force FB, which is perpendicular to the force FB. The direction of FB is related to the swinging speed of the connecting rod _ . Ffp can be written as following Ffp\u00bcFB fp sgn\u00f0 _ \u00de, sgn\u00f0 _ \u00de \u00bc \u00fe1, _ 4 0 0, _ \u00bc 0 1, _ 5 0 8< : \u00f04\u00de As shown in Figure 1, a, b, and L are the geometric parameters, mpis and mpin the masses of piston and wrist-pin, respectively, and I pin pis the piston rotary iner- tia about the center of wrist-pin. Fs andMs are related to the combustion gas force, Fg the reciprocating motion of the piston, the transverse motion and the lubrication of the piston skirt\u2013liner system. In order to solve the above equation, to determine the transverse motion of the piston, one has to calculate F, Ff, M, and Mf. F, Ff, M, and Mf can be considered as the sum of two parts according to the mixed lubrication theory F \u00bc Fh \u00fe Fc \u00bc R Z Z A \u00bd p\u00f0 , y\u00de \u00fe pc\u00f0 , y\u00de cos d dy \u00f05\u00de M \u00bcMh \u00feMc \u00bc R Z Z A \u00bd p\u00f0 , y\u00de \u00fe pc\u00f0 , y\u00de \u00f0a y\u00de cos d dy \u00f06\u00de Ff \u00bc Ffh \u00fe Ffc \u00bc R Z Z A \u00f0 , y\u00de uj j u fpc\u00f0 , y\u00de d dy \u00f07\u00de Mf \u00bcMfh \u00feMfc \u00bc R Z Z A \u00f0 , y\u00de uj j u fpc\u00f0 , y\u00de \u00f0R cos Cp\u00ded dy \u00f08\u00de at UNSW Library on July 23, 2015pij" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002789_cjme.2013.03.532-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002789_cjme.2013.03.532-Figure7-1.png", + "caption": "Fig. 7. Simulated experiment platform", + "texts": [ + " Among these pinions, four pinions are linked with four motors and reducers to provide driving torque, and others are linked with four brakes and speed increasers to simulate cutter head load. The driving pinions and driven pinions are arranged by regular intermission. PLC realizes motor control and experimental data acquisition. In order to simulate all kinds of loads and observe the torque online, host computers are used to monitor, in real time, the experimental data, such as control current, revolving speed, and output torque. The simulated experimental platform is shown in Fig. 7. Traditionally, the torque of multiple pinions drives of TBM is often treated as a static or quasi-static load when cutter head driving system in TBM is designed and controlled. And torque master slave control method is widely used in existing multiple pinions driving of the TBM. The structure of the torque master slave control method is shown in Fig. 8(a). We can see that the load torque in each pinion is not in a static state because of the nonlinear coupling factors and inertia effects of driving mechanism and cutter head even though the rotational speed of cutter head is very low" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003956_s0031918x13080073-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003956_s0031918x13080073-Figure12-1.png", + "caption": "Fig. 12. Definition of segregation coefficient of solute as ratio of maximum maxCs of solute concentration in the solid phase to maximum maxCl of its concentration in the liquid phase [36]. Solid line shows general profile of concentration C found from the solution to set of equations (15)\u2013(22).", + "texts": [ + " Then, the segregation coefficient is defined as the ratio of the maximum concentration of the solute atoms in the ( )s \u03d5 ( )\u03b5 \u03d5 ( ) ( ) ( ) 1 2 ( ) 2 2 s l s s A s s Ls p p p T + \u03d5 = \u2212 \u03d5 , \u03d5 = \u2212 \u03d5 , [ ] ( ) ( ) ln , 2 2 2( ) ln ( )(1 ) 1 ln s l e m e e e RTp k p k p k k \u03b5 \u03b5 + \u03b5 \u03d5 = \u2212 \u03d5 \u03d5 = + \u03d5 \u2212 \u2212 , v ek ( )p \u03d5 2( ) (3 2 )p \u03d5 = \u03d5 \u2212 \u03d5 , 0 1 ( ) ( ) 1 ( ) (1 ) 0 dp dp p p d d \u03d5= \u03d5= \u03d5 \u03d5 \u2212 \u03d5 = \u2212 \u03d5 , | = | = . \u03d5 \u03d5 ( )g \u03d5 ( 0) ( ) ( 1) s l C C k V C C \u2261 \u03d5 \u2192 = . \u2261 \u03d5 \u2192 solid phase to its maximum concentration in the liquid (Fig. 12) in the form (23) where (24) (25) with the function h(p) defined as (26) The large velocities of heating/cooling and the sig nificant temperature gradients upon laser treatment of materials create conditions for the occurrence of rapid phase transitions [70\u201372]. The macroscopic calcula tions of thermal fields and the dynamics of melting of powder particles that were described in the preceding sections have shown that upon the melting of ultradis persed particles the melting rate reaches maximum values exceeding 10 m/s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002159_pime_proc_1969_184_061_02-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002159_pime_proc_1969_184_061_02-Figure5-1.png", + "caption": "Fig. 5. Contact ellipse for ball in nonconforming groove", + "texts": [ + " Additional notation a b b\u2019 S s, P P P b PO Major semi-axis of contact ellipse, in. Minor semi-axis of contact ellipse, in. Semiwidth of contact ellipse at y, in. Contact stress or pressure, lb/in2. Pressure at which value of pressure-viscosity expo- Absolute viscosity, reyns, lb s/in2. Equivalent viscosity in the contact ellipse, reyns, Equivalent viscosity at y = b, reyns, lb s/in2. Absolute viscosity at atmospheric pressure and tem- nent changes, lb/in2. lb s/in2. perature, reyns, lb s/in2. Fig. 5 is a representative contact ellipse for a ball in a nonconforming groove. The friction force M, about the z axis due to the section of the ellipse for y > b and y < - b is obtained by integrating the product of the friction of the elemental roller (Fig. 6) and the moment arm (10) thus, - 0 M I = 21 y d F . . . (42) t h Proc lnstn Mech Engrs 1969-70 Vol 184 P t 1 No 44 at UNIV CALIFORNIA SAN DIEGO on June 21, 2016pme.sagepub.comDownloaded from 846 COMMUNICATION - N Per cent con fo rm i t y The above can be numerically integrated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003734_acc.2015.7170966-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003734_acc.2015.7170966-Figure4-1.png", + "caption": "Fig. 4. Lateral control scenario in global X , Y coordinates and error coordinates, eY = Y \u2212 Yd and e\u03c8 = \u03c8 \u2212 \u03c8d", + "texts": [ + " Choosing y\u0307, \u03c8, \u03c8\u0307 as the states of the bicycle model, a discrete-time formulation of the dynamic equations (18), (19) is obtained by applying Euler discretization. The Euler method approximates time derivatives by forward finite differences. In the following, the resulting discrete-time state equations are augmented by error dynamics. Unlike [15], we consider error dynamics in the inertial rather than the vehicle body fixed coordinates. The components of the global vehicle velocity are X\u0307 = Vx cos(\u03c8)\u2212 y\u0307 sin(\u03c8), (22) Y\u0307 = Vx sin(\u03c8) + y\u0307 cos(\u03c8). (23) Fig. 4 illustrates the Y position error eY = Y \u2212 Yd and the yaw error e\u03c8 = \u03c8 \u2212 \u03c8d. Yd = fY (X(t)) is determined by the road shape in the case of a lane keeping algorithm or it could be the output of a vehicle path planning algorithm. It is assumed that the vehicle is primarily traveling in the global X direction. The desired yaw angle is \u03c8d = arctan(\u2202fY\u2202X ). The discrete-time error dynamics are eY (k + 1) = eY (k) + (Y (k + 1)\u2212 Y (k)) \u2212 (Yd(k + 1)\u2212 Yd(k)) , (24) e\u03c8(k + 1) = e\u03c8(k) + (\u03c8(k + 1)\u2212 \u03c8(k)) \u2212 (\u03c8d(k + 1)\u2212 \u03c8d(k)) " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000621_20080706-5-kr-1001.01254-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000621_20080706-5-kr-1001.01254-Figure1-1.png", + "caption": "Fig. 1. The miniature Eagle helicopter platform", + "texts": [ + " Numerical results as well as statistical analysis are carried out to validate the proposed technique. Section 2 details the platform used for the experimentation and data collection while section 3 deals with the formulation for system identification and neural network technique. Results and discussion are provided in section 4 and section 5 concludes the paper. In this section a brief overview of the platform and sensors used for flight experiments is provided. The platform used for experiments (Figure 1) is a Hirobo \u201cEagle\u201d 60 size radio controlled helicopter, modified to use a brushless DC motor in lieu of the standard internal combustion engine. The design incorporates a single main rotor and a single tail-rotor for anti-torque compensation. Five servo actuators: collective, throttle, aileron, elevator and tail rotor pitch are used to control the helicopter. The autopilot system consists of the MPC555 based autopilot and PC104 based flight computer (Figure 2). MPC555 based Autopilot: This unit is made in-house based on a Motorola MPC555 microcontroller" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001066_s1007-0214(09)70086-0-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001066_s1007-0214(09)70086-0-Figure1-1.png", + "caption": "Fig. 1 Working principle of Perfactory\u00ae Type III mini system", + "texts": [ + " With MSL, a complex structure can be manufactured in a single process, thus avoiding the expensive assembly procedures while dramatically improving the manufacturing reliability at the same time. The current research explores the development of a projection MSL system for the manufacturing of non-assembly complex devices with micro structures, which find wide applications in various biomedical, MEMS, and micro precision research. This research utilises a newly developed projection MSL system, the Perfactory Type III mini system (EnvisionTec, Germany). Figure 1 shows the principle of the system used. A 3-D solid model designed with CAD software is sliced into a series of 2-D layers of uniform thickness. The NC code generated from each sliced 2-D file is then executed to control a dynamic pattern generator. The generator is used to create pattern projection for exposure curing. The mask image is created by microscopically small mirrors laid out in a matrix on a semiconductor chip, known as a digital micromirror device (DMD, product of Texas Instruments)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003469_j.oceaneng.2013.01.001-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003469_j.oceaneng.2013.01.001-Figure7-1.png", + "caption": "Fig. 7. Simplified model of 4-legged robot system.", + "texts": [ + " In addition, we know that the magnitude of translational accelerations along x, y and z axes in positive area of z axis are bigger than ones in the other area as shown in Fig. 9(c, d). From these results and the physical intuition, we guess that the proposed method has high validity in the aspect of acceleration analysis considering a robot\u2019s posture. Next, we have performed the simulation considering differential environment for analysis of the influence of underwater environment on multi-legged robot. To compare the simulation results of ground and underwater case easily, this simulation has been performed with the same example as shown in Fig. 7. For the simulation of underwater case, we assume that seabed reaction force rate is 90%, tidal current velocity is 1 knots, drag coefficients are CBP\u00bc2 and CD\u00bc1.1, and buoyancy is 20% of gravity as shown in Table 4. The rest of parameters are the same of example case I. In this case, we caneasily expect several results from the parameters related to underwater environment: (1) all acceleration bounds are smaller than ones of example case I (due to reduction of seabed reaction forces), (2) the translational acceleration bound along positive Z axis is shifted to the negative direction of X axis (due to the tidal current), (3) the minimum acceleration along Z axis is larger than one of ground case (due to the buoyancy)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000796_rd.186.0534-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000796_rd.186.0534-Figure9-1.png", + "caption": "Figure 9 Schematic for slider dynamics. I I is the moment of inertia; \u00ab; Kf3' spring stiffness; 0', f3, pitch X a'rid roll angles, re spectively; N, F, W, forces on slider.", + "texts": [ + " As we have seen previously, very good agreement in the high speed limit exists with experimental measure ments, and it appears that the steady state Reynolds equation sufficiently describes the physical situation in this limiting case. Although the solution of the steady state equation is appropriate for large velocities and very smooth sur faces, in the case of appreciable slider dynamics the time-dependent Reynolds equation and the dynamic equations of motion [11] of the slider must be used to describe the physical situation. The dynamic equations of the slider shown in Fig. 9 assume then the following form: It seems justifiable to postulate that Eqs. (4) and (5) are valid not only for the description of the dynamic behavior of the slider bearing, but also for the descrip tion of the slider motion between intermittent contacts. The equations, however, are not valid for the intermit tent contact itself. To describe contacts we may proceed in the following way. We first calculate the dynamic flying height from Eqs. (4) and (5) by simultaneous numerical solution. Whenever the calculated spacing is less than the sum of the surface roughness of the disk and the slider, an intermittent contact is assumed by in cluding the effect of a normal impact force N* (r) in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001975_s12555-010-0301-x-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001975_s12555-010-0301-x-Figure1-1.png", + "caption": "Fig. 1. A ball and beam system.", + "texts": [ + " Wen Yu is with the Departamento de Control Autom\u00e1tico, CINVESTAV-IPN, Av.IPN #2508, Mexico D.F., 07360, Mexico (e-mail: yuw@ctrl.cinvestav.mx). * Corresponding author. Xiaoou Li and Wen Yu 492 compensation is also proven. Experimental tests are carried out to show the effectiveness of the algorithms proposed in this paper. 2. BALL AND BEAM MODEL AND PD CONTROL The control objective of ball and beam systems is to turn the angle of gear ,\u03b8 and the angle of the beam ,\u03b1 such that the ball can stay in a position ,r see Fig. 1. When the angle is changed from the horizontal position, gravity causes the ball to roll along the beam. The electrical and mechanical subsystems are coupled to each other through an algebraic torque equation ( )1 , , , g m m m m b m m mK m m m U L I R I K J B K I \u03b8 \u03b8 \u03b8 \u03c4 \u03c4 = + + + = = (1) where U is input voltage, m I is armature current, m R and m L are the resistance and inductance of the armature, b K is back emf constant, \u03b8 is angular velocity. Compared to m m R I and , b K \u03b8 the term m m L I is very small, g K is gear ratio, m J is the effective moment of inertia, B m is viscous friction coefficient, m \u03c4 is the torque produced at the motor shaft, m K is torque constant of the motor", + " m b m K U K R \u03b8 \u03c4\u2212 = (2) In the absence of friction or other disturbances, the dynamics of the ball and beam system can be obtained by Lagrangian method, the mathematical model of the ball and beam system is given by ( )2 2 4 2 cos , sin 0, e J J mr mr r k r r g \u03b1 \u03b1 \u03b6 \u03b1 \u03c4 \u03b1 \u03b1 + + + + = \u2212 + = (3) where J is the moment of inertia of the beam, e J is the moment of inertia of the ball, \u03b1 is the angle of the beam, r is the position of the ball, m is mass of the ball, M is mass of the beam, L is longitude of the beam, \u03b6 is gravity of the ball and beam system, 2 , Lmgr Mg\u03b6 = + see Fig. 1. Usually the beam angle \u03b1 and motor position \u03b8 are not the same. The arc distances in the two circle are equal, i.e., .L d\u03b1 \u03b8= (4) The whole ball and beam system is (2) and (3) 2 1 2 3 2 4 ( ) 2 cos 2 , sin 0, L mr k mr r mgr Mg k U k k r r g \u03b1 \u03b1 \u03b1 \u03b1 \u03b1 \u03b1 + + + + = \u2212 \u2212 + = (5) where 1 1 , m m m g R J L k J dK K = + 2 1 ,m m K k R = + 3 m b m K KL k d R = , m m b m g R B K K K + + 4 7 . 5 k = Two types of PD control are designed for this system, see Fig. 2. The serial PD control with nonlinear compensation has the following form ( ) ( ) , ( ) ( ), pm dm d d pb db U k k k r r k r r \u03b1 \u03b1 \u03b1 \u03b1 \u03c0 \u03b1 \u2217 \u2217 \u2217 = \u2212 + \u2212 + = \u2212 \u2212 \u2212 \u2212 (6) where pm k and dm k are positive gains for the motor Synchronization of Ball and Beam Systems with Neural Compensation 493 control, pbk and db k are positive gains for the ball control, \u03c0 is the compensator, r d is reference position", + " Theorem 2: If the weights of neural compensator for the PD synchronization control (10) are updated as 2 2 2 \u02c6 \u02c6[ ( ) ], \u02c6 \u02c6[ ], T T i wi i i i i i i i i i i T i vi i i i i W k V x k D V x V k xW D \u03c3 \u03c3 \u03b1 \u03c6 \u03b1 \u03b1 = \u0393 + = \u0393 where , vi \u0393 wi \u0393 are learning rates, , vi \u0393 0, wi \u0393 > and the PD gains in (u1) satisfy 2 2 22 1 4 4 1 2 6 2 1 6 2 2 2 2 1 4 2 3 2 5 3 , 2 2( ) , 2 , , 2 i i i i i i i i i i i i i i i i i i i k a a a a k a k a a k k k k k a k a k > > + + > + > then the synchronization of ball and beam system is stable and the synchronization state converges 2 2 0 1 lim sup 2i T i i i iH T k x dt T \u03c3 \u03b7 \u03b7 \u2192\u221e \u2264 +\u222b from any well defined set of initial conditions, i H = 2 2 5 3 0, ,0,1, ,1 . 2 i i i i i k diag k a k L + \u2212 Synchronization of Ball and Beam Systems with Neural Compensation 495 5 EXPERIMENTAL CASE STUDY The synchronization for ball and beam is carried out on two types of systems, one is made by Quanser [16], the other is made by the Balance Control Inc., see Fig. 4. The physics parameters (see Fig. 1) for Ball and beam 1 are 1 1 1 1 1 60 , 0.12 , 0.06 , 16, 0.3.m m m L cm M kg m kg RL d R K = = = = = + The physics parameters for Ball and beam 2 are 2 2 2 2 2 70 , 0.2 , 0.85 , 1, 0.7.m m m L cm M kg m kg RL d R K = = = = = + The nonlinear compensator is [ ]2 1 2 2 3 2 2 1 2 2 2 sin 0 1 [( ) cos 2 ( ) 2 ]cos 1 cos 0, 2 r k a g r r k L mgr Mg k a r k k mr rr L mgr Mg a r k \u03b1 \u03b1 \u03b1 \u03b1 \u03b1 \u03c0 \u03b1 \u03b1 \u03b1 \u03b1 \u03b1 + \u2212 + \u2260 + + \u2212 = \u2212 + \u2212 + \u2212 = where 1 ,pm pba k k= 2 ,pm db dm pba k k k k= + 3 , dm db a k k= 4 , pm a k= 5 , dm a k= they satisfy 2 2 2 4 2 1 2 5 3 2 2 1 , " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000281_1.2777476-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000281_1.2777476-Figure3-1.png", + "caption": "FIG. 3. Schematics of adhesive microcontact.", + "texts": [ + " To avoid numerical calculations based on a specific analytical solution, Maugis suggested the MD adhesion model utilizing the Dugdale approximation, which assumes a state of constant adhesive stress over some length at the gap between a contacting sphere and a flat. The constant adhesive stress is defined by the work of adhesion, which equals the work expended to move the two half-spaces from the equilibrium distance to infinity. As shown in Fig. 2, the MD approximation can also be expressed as w = 0h0, 1 where w is the adhesive work, also referred as surface energy, h0 is the effective range of the Dugdale stress in the vertical direction, and 0 is the Dugdale stress. Figure 3 is the physical description of the adhesive contact of microasperity in the presence of the intermolecular adhesive interactions. Combing the fracture mechanics solu- tion for a symmetric crack subjected to Dugdale-type traction, Maugis obtained the Maugis-Dugdale theory when in elastic adhesion, which is 1 = A2 2 m2 \u2212 1 + m2 \u2212 2 arctan m2 \u2212 1 + 4 2A 3 m2 \u2212 1 arctan m2 \u2212 1 \u2212 m + 1 , 2 P\u0304 = A3 \u2212 A2 m2 \u2212 1 + m2 arctan m2 \u2212 1 , 3 = A2 \u2212 4 3 A m2 \u2212 1, 4 where P\u0304= Pe / wR, = / 2w2R /K2 1/3, A=a / wR2 /K 1/3, =2 0 / wK2 /R 1/3, and m=c /a are the dimensionless load, approach, contact radius, adhesive contact radius, transition parameter, and cohesive to contact radius ratio, respectively, a is the radius of contact area, c denotes the radius of cohesive zone, and x is the distance to the center line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002609_icca.2011.6138036-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002609_icca.2011.6138036-Figure2-1.png", + "caption": "Fig. 2. UAV references", + "texts": [ + " It was shown in [19] that if there exists a positivedefinite C 1 function \u03a0(x) locally defined in a neighborhood of the origin in R n that satisfies H (x,\u03a0T x ,\u03b11(x,\u03a0 T x ),\u03b12(x,\u03a0 T x )) = 0, or, more explicitly, \u03a0x f (x)+ 1 2 hT 1 h1 + 1 2 \u03a0x ( 1 \u03b32 g1gT 1 \u2212 g2gT 2 ) \u03a0T x = 0, (3) where \u03a0x denotes the Jacobian matrix of \u03a0(x), then the control law u =\u2212gT 2 (x)\u03a0 T x (x) (4) solves the nonlinear H\u221e state feedback problem. The H\u221e controller will be designed for the a complete nonlinear model of a fixed wing UAV. The complete nonlinear model for a fixed wing UAV developed by [18] has the state vector x = [U V W P Q R \u03c6 \u03b8 \u03c8 h \u03c9 ]T , (5) where U , V , W are the velocities in m/s, P, Q, R are the angular speeds in rad/s, \u03c6 , \u03b8 , \u03c8 the Euler angles in rad, h the altitude in m and \u03c9 the angular speed of the propeller in rad/s according to Fig. 2. This type of aircraft has as control inputs the propeller\u2019s motor throttle (\u03b4T , in percentage), and the deflection angles of the four control surfaces: elevator (\u03b4e), two ailerons (\u03b4a), and rudder (\u03b4r). The two ailerons constitute one control input (\u03b4a) since they always act in a coordinated manner. The control input vector is then u = [ \u03b4e \u03b4a \u03b4r \u03b4T ]T . (6) The space state model is built using the force equations, the kinematics equations, the moment equations, the Z-axis navigation equation and the propeller\u2019s dynamic" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001559_s00034-009-9126-3-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001559_s00034-009-9126-3-Figure3-1.png", + "caption": "Fig. 3 Full-vehicle suspension system", + "texts": [ + " (6) According to (41), the desired feedback linearization controller ufeedback can be constructed such that the uniform ultimate bounded stability is guaranteed. That is, the system dynamics enter a neighborhood of zero state and remain within it thereafter. (7) Finally, the conventional fuzzy control ufuzzy is immediately applied to improve the convergence rate of the tracking error dynamics. 3 Illustrative Example A full-car model active suspension system is introduced in this section. The system entails four suspension mechanisms and four control units as shown in Fig. 3 [8, 10, 11, 41], and is represented as a nonlinear seven-degree-of-freedom (7-DOF). All suspensions consist of a spring, a damper and an actuator to generate a pushing force between the body and the axle. The sprung mass is assumed to be a rigid body and has freedom of motion in the vertical (heave), pitch and roll directions, while the unsprung masses are free to bounce vertically with respect to the sprung mass. The suspension between the sprung mass and the unsprung masses are modeled as linear viscous dampers and spring elements. The tire is assumed to contact the surface of the road when the automobile is travelling; and is regarded as a spring with a mass. From Fig. 3, the displacements of the sprung mass are given in the following: \u2022 Front-left wheel zs1 = z + a sin \u03b8 + d sin\u03d5 (93) \u2022 Front-right wheel zs2 = z + a sin \u03b8 \u2212 c sin\u03d5 (94) \u2022 Rear-left wheel zs3 = z \u2212 b sin \u03b8 + d sin\u03d5 (95) \u2022 Rear-right wheel zs4 = z \u2212 b sin \u03b8 \u2212 c sin\u03d5 (96) Here zs1 is the front-left body displacement, zs2 is the front-right body displacement, zs3 is the rear-left body displacement, zs4 is the rear-right body displacement, a and b are the distances from front wheels and the rear wheels to the center of mass of the sprung mass, c and d are the distances from the right wheels and the left wheels to the center of mass of the sprung mass, and z, \u03b8 , \u03d5 are the vertical displacement, the pitch angular displacement and the roll angular displacement of the center of mass of the sprung mass" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001497_j.matdes.2010.02.029-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001497_j.matdes.2010.02.029-Figure1-1.png", + "caption": "Fig. 1. Skywriting scanning option [16].", + "texts": [ + " To avoid this situation and keep acceleration and retardation phase out of the layer geometry, skywriting option is provided in commercial SLS machines. If skywriting option is selected, then the mirror is accelerated already before the start of the part so that it has reached the desired speed before the beginning gy with skywriting [16]. Fig. 5. Up-down skin option [16]. of exposure. Laser is switched on at the start of the part. Similarly retardation phase begins at the part end where the laser is switched off as shown in Fig. 1. The sorted strategy (Fig. 2) searches the shortest exposure path across the part, whereas the unsorted strategy (Fig. 3) moves across the part in the easiest way. The sorted exposure type is usu- Fig. 6. Solid model has been divided int ally faster than the unsorted type but it may lead to gaps or slight recesses between first and second exposure phase as shown in Fig. 2. Four choices for hatch pattern selection are generally available, i.e., along X, along Y, both in X and Y or alternating in X and Y as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000826_j.euromechsol.2009.07.005-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000826_j.euromechsol.2009.07.005-Figure1-1.png", + "caption": "Fig. 1. General representative of {X 1(u)}{X 2(v)} generators of X\u2013X motion.", + "texts": [ + " We emphasize once more that these submanifolds are not equal. The products with the type denoted {X 2(u)} have two factors chosen from {H(N, u, p)} for any point N and any pitch p and {T(s)}, s being any unit vector. However, the product {T(s1)}{T(s2)} is a subset of {T} and has to be rejected. Hence, there are (i) {H(N1, u, p1)}{H(N2, u, p2)}, in which the point N2 must not lie on the line (N1, u); (ii) {H(N1, u, p1)}{T(s)} and all other factor permutations. One representative chain of this family is shown in Fig. 1. Moreover, Table 3 symbolically shows all the members of this family of X\u2013X motion generators and the corresponding chains are graphically displayed in Fig. 2 for possible applications more easily. We recall that a revolute R pair is the special case of an H pair with a zero pitch, and therefore the replacement of one or several Hs by Rs in the above-tabulated chains leads to special cases of the generators. Clearly, another irreducible representation of the product {X(u)}{X(v)} is {X 2(u)}{X 1(v)}" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002852_s0022112010006075-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002852_s0022112010006075-Figure2-1.png", + "caption": "Figure 2. A schematic diagram of a section of the helical ring depicting its one-dimensional centreline \u0393 and surface \u03a3 .", + "texts": [ + " On the basis of these definitions, we set s \u2032 = s/ , R\u2032 = R/ and \u03bb\u2032 = \u03bb/ to get the dimensionless flagellum centreline X\u2032(s \u2032, t \u2032) = [ 1/2 \u2212 R\u2032 sin ( 2\u03c0s \u2032 \u03bb\u2032 \u2212 2\u03c0t \u2032 )] cos(2s \u2032), Y \u2032(s \u2032, t \u2032) = [ 1/2 \u2212 R\u2032 sin ( 2\u03c0s \u2032 \u03bb\u2032 \u2212 2\u03c0t \u2032 )] sin(2s \u2032), Z\u2032(s \u2032, t \u2032) = R\u2032 cos ( 2\u03c0s \u03bb \u2212 2\u03c0t \u2032 ) , \u23ab\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ad (2.5) so that the dimensionless wave speed is w\u2032 s = \u03bb\u2032. We also define np = \u03c0/\u03bb\u2032 as the integral number of pitches around the ring. For the remainder of the paper, we use the dimensionless formulation of the problem and drop the primes on the variables. We point out that when using slender-body theory (Lighthill 1976), these forces are supported only along the one-dimensional centreline \u0393 of the helical ring (see figure 2); however, when using regularized Stokeslets (Cortez 2001; Cortez, Fauci & Medovikov 2005), the forces are distributed on the surface \u03a3 of the helical ring. In either case, the Stokes equations imply that the distribution of forces is linearly related to the distribution of velocities. 2.1. Instantaneous fluid velocity computations In the static computations of instantaneous flows, we begin with an equation that encodes the linear relationship between the fluid velocity and the forces exerted by the flagellum" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003312_j.jsv.2013.05.022-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003312_j.jsv.2013.05.022-Figure3-1.png", + "caption": "Fig. 3. Open-circuit magnetic field distributions, where (a) 1-magnet/15-slot, (b) 8-magnet/1-slot, and (c) 8-magnet/15-slot.", + "texts": [ + " In this sense, the theoretical predictions can be equivalently simulated merely based on the two cases, including the GCD equal to unity, greater than unity, or equal to magnet number. This section examines the synthesis effect of the fields from each magnet in the twomotors. Imaginary motors with 1-magnet/ 15-slot, 8-magnet/1-slot, and 8-magnet/15-slot are presented for the odd-slot motor, and those with 1-magnet/12-slot, 8-magnet/ 1-slot, and 8-magnet/12-slot are also given for the even-slot motor. Open-circuit magnetic field distributions are presented. Fig. 3 illustrates the magnetic field distributions. The results imply that fields in the 1-magnet/15-slot and 8-magnet/ 1-slot motors exhibit weak symmetry (Fig. 3(a) and (b)). Namely, the magnetic fields cannot be balanced, and thus UMF is produced. By contrast, there exists a relatively more symmetry for the 8-magnet/15-slot motor, as shown in Fig. 3(c). The results reflect certain force characteristics for different magnet/slot combinations. It is easy to imagine that the more the magnet/slot number is, the more symmetry the flux distribution is, and thus the more balance the magnetic force can be. The radial and tangential magnetic flux densities are presented to further clarify the effect of magnet/slot combination on magnetic forces, as shown in Fig. 4. The results show that the waveforms in the 1-magnet/15-slot and 8-magnet/1-slot motors exhibit weak symmetry, which repeat themselves after a complete rotation as the magnet number is prime to slot number", + " In contrast, the 24th order is definitely excited because it is the equivalent 3rd order of the magnet-frequency or the 2nd order of the slot-frequency forces, where the factors are km\u00bcmod(lmNm/Ns)\u00bcmod(3 8/12)\u00bc0 and ks\u00bcmod(lsNs/Nm)\u00bcmod(2 12/8)\u00bc0, respectively. 3.3.3.1. SHAFT-FREQUENCY EXCITATION ANALYSIS. When considering the shaft frequency force, quite a different situation arises, thought the occurrence or suppression behaviors of the magnetic forces can also be addressed in a similarly way. The 1-magnet/15-slot motor in Fig. 3 and 1-magnet/12-slot motor in Fig. 5 are equivalent to rotors with moving eccentricity, where as the 8-magnet/1-slot motors in Figs. 3 and 5 can model the stators with eccentricity, where the eccentricity is movable with the rotating stator for the odd-slot motor but unmoved for the even-slot motor. 3.3.3.2. HARMONIC FORCE IN THE 8-MAGNET/15-SLOT MOTOR. Physically, the 1st, 14th, 29th, and 31st orders in Fig. 7, (c) and (d) are all related to the shaft-frequency for the 1-magnet/15-slot motor. These are remarkable because they are in essence the 0th, 15th, and 30th harmonics, where the 0th harmonic corresponds to the direct component in the force on the magnet causing UMF and CT, and the rest two have the factors km0 \u00bcmod(lm/Ns)\u00bc{mod(15/15), mod(30/15)}\u00bc{0, 0}, also implying excited UMF and CT" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000009_12.723300-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000009_12.723300-Figure3-1.png", + "caption": "Fig. 3: The geometry for parallel guidance", + "texts": [ + " It is obvious that cooperation between the pursuers is required. Then our task is to coordinate all the pursuers such their angle distribution relative to the evader is invariant and at the same time some or all the pursuers approach the evader until capture. We here propose to use the parallel guidance law, which is a popular method in missile guidance and control18, in the pursuers\u2019 phase control to implement the parallel relationship of current LOS EPi and the next LOS E\u2019P\u2019i ' 'EP E P ( 1, 2, , )i i i n= L (8) As illustrated by Figure 3, after players move from their current position E, P1 and P2 to next position E\u2019, P\u20191 and P\u20192, we must have that ' ' 1 1EP E P and ' 'EP E Pn n . We have the following theorem. Theorem 1: There are n pursuers (with identical maximal speed u) angle-evenly distributed around the evader (with maximal speed v), if 1v u > (9) min 2 2arcsin n n u v \u03c0 + \u23a1 \u23a4 \u23a2 \u23a5 \u23a2 \u23a5\u2265 = \u239b \u239e\u23a2 \u23a5 \u239c \u239f\u23a2 \u23a5\u239d \u23a0\u23a3 \u23a6 (10) Proc. of SPIE Vol. 6578 657811-4 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/20/2016 Terms of Use: http://spiedigitallibrary" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002257_j.jsv.2011.06.018-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002257_j.jsv.2011.06.018-Figure3-1.png", + "caption": "Fig. 3. Model description: (a) the sliding equilibrium and (b) the perturbed state at the sliding equilibrium.", + "texts": [ + " The ball is in contact with the hemispherical shell-liner inserted into the rigid socket where the liner is modeled as contact stiffness (kc). The pre-axial load (No) and the rotation of the beam generate normal and friction stresses on the contact surface between the ball and the liner as shown in Fig. 2. In this paper, the dynamic stability criteria at the sliding state are investigated for the prediction of squeak propensity. Therefore, the sliding configuration is found under the action of pre-normal and friction loading and then the perturbation at the sliding equilibrium is described as shown in Fig. 3. According to the above scenario, the sliding equilibrium is to be determined first. Under the sliding state, the center of the ball is located at \u2018\u2018A\u2019\u2019 and the corresponding contact location can be described as req c \u00bc xA\u00feRer , (1) xA XAio\u00feYAjo\u00feZAko, (2) where xA is the position vector of \u2018\u2018A\u2019\u2019 at equilibrium and er is the radial direction vector of the spherical coordinates as shown in Fig. 4. The corresponding velocity vector at the contact location is given by veq c \u00bcX \u00f0Rer\u00de, (3) where the time derivative of the equilibrium deflection vector becomes zero by definition ( _xA \u00bc 0) and O is the angular velocity of the local axis (X\u00bcOko)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000155_j.triboint.2007.02.018-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000155_j.triboint.2007.02.018-Figure2-1.png", + "caption": "Fig. 2. Geometric characteristics of two tapered rollers in opposite orientation and the coordinate system of the film.", + "texts": [ + " The influences of some parameters, such as the angle of rollers and the velocity of rotating, on the lubrication are analysed. The meshing teeth of a pair of helical gears are shown in Fig. 1. K1K2 is the meshing line, N1N2N3N4 the meshing plane, and B1B2B3B4 the actual meshing plane. b1 and b2 are the spiral angles between K1K2 and two axes (for helical gears, b1 b2). At the moment of meshing shown in Fig. 1, the meshing teeth of the helical gears can be transformed approximately to two tapered rollers (a and b) in opposite orientation as shown in Fig. 2. The axes of rollers a and b shown in Fig. 2 correspond to lines N4N3 and N1N2, respectively in Fig. 1, and the contact line of two tapered ARTICLE IN PRESS P. Yang, P. Yang / Tribology International 40 (2007) 1627\u20131637 1629 rollers shown in Fig. 2 corresponds to line K1K2 in Fig. 1. In Fig. 2, ra and rb are the section radii on the middle section of tapered rollers a and b, respectively, ral and rbl are the section radii at the large ends of tapered rollers, respectively, similarly, ras and rbs are the section radii at the small ends. ba and bb are the deflective angles between the common generatrix and both axes of the two rollers, respectively. Geometric relationships between the helical gears shown in Fig. 1 and the tapered rollers shown in Fig. 2 are as follows: rbl \u00bc K1N1; rbs \u00bc K2N2; ral \u00bc K2N3; ras \u00bc K1N4, ba \u00bc b1; bb \u00bc b2. A Cartesian coordinate system as shown in Fig. 2 is established for the oil film, where the x-axis is perpendicular to the sheet of paper and points to readers. The coordinate systems for solids a and b are similar to that for the film, that is, the x- and y-axes are the same as those shown in Fig. 2, and, za and zb coordinates are used for solids a and b, respectively, with the same direction as z, however, with different origins. An interesting contact area is formed between the tapered rollers as shown in Fig. 3. The ends of both rollers are modified with dub off profiles. Fig. 4 shows the shape modification of the left ends, the modified shapes of the right ends are similar to those shown in Fig. 4. Let rmal, rmbl, rmas and rmbs be the radii of modification for solids a and b in the large and small ends, respectively. The length of the generatrix is 2 l before modification (see Fig. 2), and 0.9 2 l after modification. Relations of the radii of end modification are assumed to be rmas=ras \u00bc rmal=ral, rmbs=rbs \u00bc rmbl=rbl, so only two values of rmal, rmbl, rmas and rmbs are required. In this section all symbols are defined in the Nomenclature. For the TEHL of two tapered rollers in opposite orientation, the governing equations are given as following. Since the tapered rollers are rotating around their axes, the surface velocities are varied along the contact line. For the rollers in Fig. 2, the velocities of surfaces a and b can be expressed as ua \u00bc oa\u00f0ra \u00fe y sin ba\u00de; va \u00bc oax; ( (1) ub \u00bc ob\u00f0rb y sin bb\u00de; vb \u00bc obx: ( (2) Note that the shape modification of the ends of rollers was not considered when Eqs. (1) and (2) were derived, because the gaps between the modified surfaces are very large, so that the pressures there are very low. The contact area shown in Fig. 3 is in fact very narrow, i.e., the changing range of x is much smaller compared with that of y and the sizes of ra and rb, hence the absolute values of va and vb given by Eqs", + " (3), the boundary and cavitation conditions are given by p\u00f0xin; y\u00de \u00bc p\u00f0xout; y\u00de \u00bc p\u00f0x; yin\u00de \u00bc p\u00f0x; yout\u00de \u00bc 0, p\u00f0x; y\u00deX0\u00f0xinoxoxout; yinoyoyout\u00de. The above conditions mean that, during the process of pressure relaxation, once a negative nodal pressure was obtained, it was immediately forced to be zero. In this way the cavitation boundary conditions (where both pressure and pressure gradient are zero) at the unknown edge of the fluid film in the outlet region were satisfied automatically. For the two tapered rollers in Fig. 2, expression of the gap between solid surfaces can be derived through a geometric analysis. The film thickness h can be written as h \u00bc h0 \u00fe h1 \u00fe h2, (4) where h0 is a constant depends on the applied load, and h2 is the sum of normal deformations of contact surfaces due to pressure: h2 \u00bc 2 pE0 ZZ p\u00f0x0; y0\u00deffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x x0\u00de2 \u00fe \u00f0y y0\u00de2 q dx0 dy0. In Eq. (4), h1 is the geometric gap and can be written as h1 \u00bc hb ha, where hb \u00bc d2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi d22 d1x2 q \u00fe f 2 \u00f0y\u00fe 0:9l\u00de2 2rmbl \u00fe f 1 y 0:9l\u00f0 \u00de 2 2rmbs , d1 \u00bc cos2 bb cos 2bb , d2 \u00bc y sin bb cos bb rb cos bb cos 2bb " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002973_s0373463313000556-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002973_s0373463313000556-Figure3-1.png", + "caption": "Figure 3. Thruster configuration.", + "texts": [ + " The reduced position state vector \u03b7=[x y z \u03c8]T and velocity state vector q=[u w r]T. The dynamics modelling of the AUV can also be simplified as follows (Lapierre and Jouvencel, 2008; Edin and Geoff, 2004; Zhu et al., 2011): (m\u2212 Xu\u0307)u\u0307+ Xuu+ Xuuu|u| = \u03c4X (4) (m\u2212 Zw\u0307)w\u0307+ Zww+ Zwww|w| = \u03c4Z (5) (IZ \u2212Nr\u0307)r\u0307+Nrr+Nrrr|r = \u03c4N (6) The AUV has four thrusters; two thrusters are in the horizontal plane and the other two are in vertical plane. Figure 2 shows its structure. A brief sketch of the vehicle\u2019s thruster configuration is shown in Figure 3. The dynamic equation of the AUV can be written by Equation (7), based on the thrusters\u2019 distribution. We can deduce the relationship between the normalized form of torque and the normalized form of thruster forces as follows: \u03c4 = \u03c4X \u03c4Z \u03c4N = K K 0 0 0 0 K K KR \u2212KR 0 0 T1 T2 T3 T4 (7) \u03c4Zm = 2KTm (8) \u03c4Zm = 2KTm (9) \u03c4Nm = 2KRTm (10) \u03c4X \u03c4Xm \u03c4Z \u03c4Zm \u03c4N \u03c4Nm = 1 2 1 2 0 0 0 0 1 2 1 2 1 2 \u2212 1 2 0 0 T1 Tm T2 Tm T3 Tm T4 Tm \u21d4 \u03c4\u0304X \u03c4\u0304Z \u03c4\u0304N = B \u00b7 T1 T2 T3 T4 \u21d4 \u03c4\u0304 = B \u00b7 T (11) where \u03c4 are the values of the surge and heave forces and the yaw torque respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001853_j.neuint.2009.10.004-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001853_j.neuint.2009.10.004-Figure4-1.png", + "caption": "Fig. 4. Time course of [2-14C]serotonin uptake by synaptosomes isolated from brains of 1\u2014control rats; 2\u2014diabetic rats; 3\u2014diabetic rats treated with MNA in dose 100 mg/kg b.w., given with drinking water for 5 weeks. Synaptosomes were preincubated for 5 min at 37 8C and assayed for the influx of [2-14C]serotonin as described in Section 2. Iproniazide (monoamine oxidase inhibitor, 10 mM) was presented in standard assay solution to minimize serotonin catabolism. Nonspecific [2-14C]serotonin accumulation was defined in the presence of 10 mM citalopram. Values for [2-14C]serotonin uptake are expressed as pmol/mg of protein. Data represent means S.E.M. for three independent experiments each performed in quadruplicate. p < 0.05 between control and diabetic groups and between untreated and treated diabetic groups.", + "texts": [ + " Enhanced glucose flux through polyol (sorbitol) pathway due to glucose over-consumption may be implicated as one of the important pathogenic mechanisms underlying brain failures observed under diabetic conditions (Oates, 2008). As shown in Table 1, diabetes was associated with an elevation of brain sorbitol level compared with control animals by over 62.1%. Diabetic rats treated with MNA exhibited 33.1% reduction in brain sorbitol content. To determine whether diabetes-induced disturbances in ion transport systems and synaptosomal membrane potential decline contributed to the alterations of synaptic function, [2-14C]serotonin uptake and release experiments were carried out. Data presented in Fig. 4 indicate that synaptosomes isolated from diabetic rats reduced total [2-14C]serotonin uptake by 37% compared to controls. MNA treatment provided 15% more efficient uptake of the neurotransmitter than it is seen in diabetes. Spontaneous [2-14C]serotonin release increased in diabetes up to 13.24 0.78 vs. 9.35 0.51 in control and MNA administration to diabetic rats slightly lowered mediator release to 11.03 0.67 pmol/ mg protein, p < 0.05. Maintenance of nicotinamide adenine dinucleotides on the proper level may be considered as regulating factor of ATP biosynthesis as well as ATP-dependent processes including Na+,K+ATPase activity and membrane potential generation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002160_2011-01-1548-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002160_2011-01-1548-Figure5-1.png", + "caption": "Figure 5. Shaft-bearing assembly: (a) Pinion, (b) Gear.", + "texts": [ + " | Volume 4 | Issue 2 1043 Also, represents the transpose of hl, and Fm stands for the mesh force at any given point in time, which includes both spring and mesh damping forces as follows: (17) where cm is the mesh damping ratio. The dynamic transmission error \u03b4 is defined as: (18) The nonlinear function f (\u03b4 - e) is applied to model the clearance nonlinearity inherent in the hypoid gear system, and is defined as: (19) where b is the backlash in the hypoid gear pair. A typical industrially applied hypoid geared rotor system is studied here. The gear and bearing data is listed in Table 1. The configuration of pinion and gear shaft bearing assembly is shown in Figure 5. The X direction translation stiffness for bearing 1 is given in Figure 6. SAE Int. J. Passeng. Cars - Mech. Syst. | Volume 4 | Issue 21044 Interaction between the time-varying mesh parameters and time-varying bearing stiffness is studied using a heavy load case without jump phenomenon. The light load case with time-invariant mesh parameters is applied to study the interaction between backlash nonlinearity and time-varying bearing stiffness. In this case, only the interaction between time-varying bearing stiffness and time-varying mesh model is studied" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003353_j.laa.2013.11.017-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003353_j.laa.2013.11.017-Figure1-1.png", + "caption": "Fig. 1. The vehicle model.", + "texts": [ + " Assume the ith vehicle has the dynamical model (see [54])[ x\u0307i y\u0307i ] = vi [ cos \u03b8i sin \u03b8i ] , (29) \u03b8\u0307i = vi li tan\u03d5i, (30) \u03d5\u0307i = \u03c9i, (31) J i\u03c9\u0307i = Ti, (32) Mi v\u0307i = Fi, (33) where (xi, yi) coordinates give i\u2019s location of the center of the rear axle, \u03b8i the heading angle with respect to the x-axis, \u03d5i the steering wheel\u2019s angle with respect to the longitudinal axis, vi the linear velocity of the center of the rear axle, \u03c9i the angular velocity of the steering wheels, li the length between the steering wheels and the rear wheels, J i the moment of inertia, Mi the mass, Ti the external torque and Fi the driving force. Fig. 1 shows some variables of vehicle i. We assume all the vehicles can only obtain the heading angles and the linear velocities through network Gt which may be switching with respect to the time. For each vehicle i, the local information used in the controllers Ti and Fi includes the state of i and the output information from i\u2019s neighbors transmitted through the network. The heading consensus problem is designing distributed controllers Ti and Fi with local information to realize |\u03b8i \u2212 \u03b8 j | \u2192 0 and |vi \u2212 v j | \u2192 0 as t \u2192 +\u221e" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002417_iros.2011.6048792-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002417_iros.2011.6048792-Figure2-1.png", + "caption": "Fig. 2. The mechanical components on HAMR3. Piezoelectric actuators provide mechanical power through the flexure-based hip joint transmission to drive the legs.", + "texts": [ + " Nominally there are twelve total degrees of freedom, which are reduced to six by a coupling scheme described below. Although the results in Section V describe HAMR3 walking straight on a flat surface, a goal for future work is to investigate a variety of gaits to enable turning, climbing, and traversing rough terrain. Therefore the actuator coupling 978-1-61284-456-5/11/$26.00 \u00a92011 IEEE 5073 scheme remains general enough to accommodate future trials as opposed to only prescribing a single gait. The mechanical components of HAMR3, illustrated in Fig. 2, are detailed below. Each leg requires two degrees of freedom for walking: lift and swing. The flexure-based spherical five-bar (SFB) mechanism in Fig. 3, which was introduced in the previous version of HAMR [8], is used to achieve this desired output. The SFB maps two decoupled drive inputs to a single end effector, in this case the leg, through a parallel mechanism that can be idealized as a ball-in socket joint sans axial rotation. Input is taken from two decoupled piezoelectric cantilever actuators through four-bar slider-crank mechanisms to the output" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003524_iccpct.2013.6528849-FigureI-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003524_iccpct.2013.6528849-FigureI-1.png", + "caption": "Figure I Elementary rotor loops [2]", + "texts": [ + " It is found that this model gives more information but requires more ca1culation time due to large number of equations required for simulation. 11. COUPLED CIRCUIT MODELLING OF INDUCTION MOTOR The model is built considering that both stator and rotor are consisting of multiple inductive circuits coupled together, and the current in each circuit is considered as an independent variable. The stator comprises of three phase concentric winding. Each of these windings is treated as a separate coil. The cage rotor consists of n bars can be described as n identical and equally spaced rotor loop. Rotor elementary model is shown in Figure I.As shown in Figure 2, each loop is formed by two adjacent rotor bars and the connecting portions of theend-rings between them. Hence, the rotor circuit has n+ 1 independent current as variables. The n rotor loop currents are coupled to each other and to the stator windings through mutual inductances. The end-ring loop does not couple with the stator windings [4], it however couples the rotor currents only through the end leakage inductance and the end-ring resistance. Figure.2 . Rotor cage Equivalent circuit [2] Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002173_304-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002173_304-Figure1-1.png", + "caption": "Figure 1. Contact between smooth spheres.", + "texts": [ + " This, however, would require rigorous mathematical treatment and, hence, the exact approach will not be adopted here. l n this paper the coupling effect will be neglected and the real contact area related to the gap between the surfaces by maintaining the shape of the deformed waviness the same throughout. 2.1. Rough spheres in contact We first consider the case of rough spheres. Before further analysis, it is necessary to establish certain relationships for smooth spheres in contact by reference to figure 1 . Consider two smooth hemispheres with radii of curvature RI and Re being loaded upon each other at a given load P. Assuming that the hemispheres deform elastically and that the stress distribution is given as the square root of the parabolic fLinction of the radial Elastic deformation of rough spheres, cylinders and annuli 1473 distance r , Hertz (1881) obtained expressions for both the radius of contact rH and the displacement (or deformation) of the surface w(r) . They are 3PR 113 Y E = (4) and If the deformed spheres shown in figure 1 are now separated by a distance d and still retain their deformed shape as shown in figure 2, the separation distance of two points on the surface, M and N, at a distance r from the centre line may then be approximated by (8) r 2 2 R z= d+ -+ w(v) - w(0) where w is a displacement function. It is interesting to observe in equation (8) that a parameter d, denoting the minimum separation distance, has been included to take into account the roughness of the spheres. Tt will be shown in the following section that such a parameter will depend on the surface topography and material properties as well as the applied load" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure4-1.png", + "caption": "Fig. 4 Generation by crown gear: \u201ea\u2026 pitch cones and \u201eb\u2026 intersection line L of generating surface with crown gear pitch cone", + "texts": [ + " \u2022 blade angle g \u2022 mean cutter radius Ru \u2022 point width Pw \u2022 radial setting Sr2 \u2022 basic cradle angle q2 \u2022 blank offset Em2 \u2022 sliding base XB2 \u2022 machine center to back XD2 \u2022 machine root angle m2 \u2022 ratio of gear roll m2c2 = 2 / c2 No tilt angle is considered in the derivations of the basic machine-tool settings although application of such an angle is commonly used in industrial application to reduce the number of standard cutter blades 20 . 3 Virtual Generation by a Crown Gear Simultaneous meshing of a bevel pinion and a bevel gear with a crown gear is similar to the simultaneous meshing of a spur pinion and a spur gear with a rack. Figure 4 a shows the pitch cones of the bevel pinion and the bevel gear with pitch angles 1 and 2. The pitch cones rotate with velocities 1 and 2 around intersected axes z1 and z2. Axis z1 and z2 make an angle . The instantaneous axis of rotation is the line of action of the relative angular velocity, line OI. Plane is a tangent plane to the pitch cones. Plane , limited with the circle of radius OI, may be considered as a particular case of a pitch cone having an apex angle =90 deg and an outer cone distance equal to OI", + " 3 Gear machine-tool settings: \u201ea\u2026 geometry of the blade f head cutter and \u201eb\u2026 installation of coordinate systems aplied for gear generation ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 03/17/201 OCTOBER 2010, Vol. 132 / 101002-3 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use c s i O a i e a g f b s M F b b g 4 m s c s i H i a t t g i b a g r e p r F t f a 1 Downloaded Fr utter in its cutting motion. The intersection line L of generating urface and pitch plane is a circular arc with radius rc, as shown n Fig. 4 b . Here, point M is the intersection point of lines L and I. The tangent to line L at M makes an angle with line OI. Such ngle is the spiral angle . The line L is approximated to an nvolute curve for the Palloid system of Klingelnberg and to an xtended epicycloid for the Cyclo-Palloid system of Klingelnberg nd Oerlikon system. Figure 5 shows the location of system Sg rigidly connected to enerating surface g and intersection line L on the crown gear or the right-hand gear Fig. 5 a and left-hand gear Fig", + " Derivation of Gear Machine-Tool Settings The following subsections show the derivations of gear achine-tool settings for the case of a spiral bevel gear drive with tandard taper. The obtained relations are then extended for the ase of uniform and duplex taper. 4.1 Derivation for Standard Type of Taper. Figure 6 a hows the pinion and gear root lines, which are not parallel but ntersect each other at the common apex of pitch cones, point O. ere, axis zf coincides with axis zcg and axis xf coincides with the nstantaneous axis of rotation, line OI see Fig. 4 a . Since the xis of rotation of the generating surface must be perpendicular to he closer generatrix of the root cone, different axes of rotation of he generating surfaces are needed for generation of pinion and ear. Axis zP is perpendicular to the gear root line whereas axis zF s perpendicular to the pinion root line. Both axes do not coincide ut make an angle 1+ 2, wherein i i=1,2 is the dedendum ngle. Different generating surfaces P and F are applied for eneration of gear tooth surface 2 and pinion tooth surface 1, espectively", + " Condition of parallelism of vectors P2 and OP is given by \u2212 2 cos R2 Am cos 2 = P \u2212 2 sin R2 \u2212 Am sin 2 7 hich yields P 2 = sin 2 8 cos 2 ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 03/17/201 Such relation is satisfied since point P belongs to the instantaneous axis of rotation by cutting of surfaces P and 2. Since the instantaneous axis of rotation by cutting of surfaces P and 2 coincides with the instantaneous axis of rotation by meshing of surfaces 1 and 2, this means that surfaces P, 2, and 1 will be in tangency at reference point P. Point P would coincide with point M defined in Fig. 4 b , 5 a , and 5 b . Figure 8 shows the generating surfaces corresponding to the convex and concave sides. The generating surface is a cone obtained by rotation of the edge of the blade. This rotation is not related with process of generation but with the required cutting velocity. The surface parameters are u , , wherein u is the profile parameter and is the longitudinal parameter. The generating surface is defined basically with two magnitudes see Fig. 8 , the blade angle gi i=1,2 and the cutter point radius Rgi i=1,2 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000514_978-3-540-30301-5_28-Figure27.6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000514_978-3-540-30301-5_28-Figure27.6-1.png", + "caption": "Fig. 27.6 The moment-labeling representation of the composite wrench cone of contact forces the pipe can apply to the pipe clamp. In this figure, the wrench wext applied to the pipe clamp can be resisted by forces within the composite wrench cone", + "texts": [ + " This means that the right- and left-hand sides of (27.14) are zero. With this assumption, we often solve for part velocities rather than accelerations. These velocities must be consistent with the kinematic constraints and force constraints, and forces acting on the parts must always sum to zero. While rigid-body mechanics problems with friction are usually solved using computational tools for CPs and LCSPs, equivalent graphical methods can be used for some planar problems to assist the intuition. As a simple example, consider the pipe clamp in Fig. 27.6 [27.14]. Under the external wrench wext, does the clamp slide down the pipe, or does it remain fixed in place? The figure uses moment labeling to represent the composite wrench cone of contact forces that can act on the clamp from the pipe. The wrench wext acting on the clamp can be exactly balanced by a wrench in the composite wrench cone. This is evident from the fact that the wrench opposing wext passes around the + labeled region in a counterclockwise sense, meaning that it is contained in the contact wrench cone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001741_j.commatsci.2009.02.007-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001741_j.commatsci.2009.02.007-Figure3-1.png", + "caption": "Fig. 3. Finite element model of the WBK unit cell with enlarged image of a brazed joint in the SUS304 stainless steel WBK specimen.", + "texts": [ + " Configurations of (a) two-dimensional Kagome truss woven by three directiona surrounded by the dashed line are, respectively, converted into (b) tetrahedrons and (c) wires (WBK). Finite element method was employed to investigate the mechanical performance of the bulk WBK, which is consisted of many identical unit cells. By imposing periodic boundary conditions, the analysis on the bulk WBK was done only on the single unit cell. The unit cell was modeled by using a commercial graphics code, 3-D PATRAN 2005 as shown in Fig. 3. The finite element analysis was performed by using ABAQUS version 6.5. The cross section of the wire was modeled by 20 quadratic brick elements (C3D20 element of ABAQUS) and the filler metal brazed at the cross points (see the figure) was modeled by 5802 quadratic tetrahedron elements (C3D10 element of ABAQUS). A variety of different heights of the filler metal from 0 to 6 mm was used to investigate the size effect of the brazed part (see Fig. 4). The unit cell model for the analysis was composed of 145,992 elements and 331,174 nodes in total", + " If a sample of WBK is composed of infinite number of unit cells, the effect of the outer surfaces on the bulk material properties can be ignored and the mechanical behavior can be estimated by analyzing the inner material. Suppose that all the cells in the inner material deform uniformly under the external force acting on the bulk sample. The deformation of the cells should be compatible to each other. Namely, the deformed shape of a unit cell must be matched with those of the neighbors. To apply the periodic boundary conditions by MPC, a reference nodal point is defined at an arbitrary position around the unit cell model (see Fig. 3). The constraint equations of the unit cells for the periodic boundary conditions are represented by ~ujji ~uj j0 i \u00bc ~ui\u00f0Ci\u00de; \u00f01\u00de where~ujji and~ujj 0 i are the displacement vectors of two surfaces periodic in i-th orientation (i = 1, 2, 3) and j and j 0 denote the index of nodes on the periodic surfaces (j, j 0 = 1, 2, ..., n). The three periodic orientations (i = 1, 2, 3) in the current problem are corresponding to the direction of x, y, z-axis, respectively. The displacement vectors~ui\u00f0Ci\u00de are defined at the reference points Ci in the corresponding periodic orientation. Every pair of nodes denoted by j and j 0 periodic in i-th orientation are displaced with constant vectors of~ui\u00f0Ci\u00de. Then a displacement field, ~ui\u00f0Ci\u00de is applied on the reference point Ci. For example, if a normal load is applied in z-axis, the displacement field ~u3\u00f0C3\u00de \u00bc uz\u0302 is applied on the reference point C3. All the other displacement vectors on the reference points are zeroes, i.e., ~u1\u00f0C1\u00de \u00bc~u2\u00f0C2\u00de. Fig. 5 shows the deformed shape of the original model in Fig. 3 under a compressive load. Note that the repeated pattern of the deformed unit cells is equivalent to the bulk WBK in Fig. 5b. The stress\u2013strain curve of the single unit cell WBK was obtained as in Fig. 6. Experimental result of the specimen of a double layered WBK and theoretical prediction are also presented in the figure for comparison. The theoretical prediction is based upon the assumption of an ideal Kagome structure [12] in which the WBK is composed of straight struts with ball joint ends" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001114_0951192x.2010.528033-Figure20-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001114_0951192x.2010.528033-Figure20-1.png", + "caption": "Figure 20. FMS layout.", + "texts": [ + " This table shows the convergence of the OF of each robot to ensure that the solution tends to move towards the same value to have an optimised solution. The optimum placement of each of the robots with respect to the work cell is shown in Figure 19. A simple FMS is used for the evaluation of the visualisation effectiveness using the UNIGRAPHICS software package. The FMS consists of an input supply of work pieces, a centre lathe, two milling machines, Puma robot and a workpiece storage for the final product. A layout of the FMS given in Figure 20 is presented to illustrate the various steps of solid modelling technique of Puma robot. The solid modelling is mainly concerned with four steps: (1) creation of robot structure, (2) served locations identification, (3) determination of robot joints generalised coordinates for the served locations and (4) visualisation of the movements of Puma robot within the FMS using a run from a predefined spreadsheet. Puma robot consists of a fixed base and six moving links connected to each other by revolute joints" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003538_aim.2011.6026975-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003538_aim.2011.6026975-Figure11-1.png", + "caption": "Fig. 11. Device for Gaze Presentation", + "texts": [ + " These results indicate that the handshake request motion model can accurately represent the human hand motion for the requesting side. Fig.10 shows a handshake robot system. The robot\u2019s arm is fabricated according to the average size of a human arm [7]. It has four degrees of freedom (shoulder: two degrees, elbow, and wrist). The coordinate system of the robot is shown in Fig.10. The origin of the coordinate system is defined as a point at the center of the shoulder. Furthermore, a gaze presentation device is constructed, as shown in Fig.11. The gaze presentation device has four degrees of freedom and is installed to the upper part of the handshake robot system. A magnetic sensor (FASTRAK) is used for the measurement of the position and angle of the human hand. The handshake approaching motion model is used to calculate the desired position of the robot based on the obtained hand position of the human. Inverse kinematics is used to calculate the joint angle from the obtained desired position of the robot. The robot is controlled according to the obtained joint angle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000399_tcst.2006.883335-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000399_tcst.2006.883335-Figure3-1.png", + "caption": "Fig. 3. Quanser-DOF helicopter. [Online.] Available: http://www.quanser.com", + "texts": [ + " In this section, we provide experimental results to show the effectiveness of the proposed methodology when applied to realtime control systems. The case study regards the design of a controller to let the attitude of a laboratory 2-DOF helicopter model track a given reference trajectory, in the presence of harmonic disturbances affecting the plant input and corrupting the measurement of the pitch and yaw rates. The experiments have been performed using the Quanser 2-DOF helicopter experimental setup, shown in Fig. 3. The helicopter model is equipped with two dc motors which actuate two propellers, directed in such a way that the front (main) propeller controls the pitch motion, while the tail propeller controls the yaw motion. The inputs of the plant are the voltages applied to the two motors, and the pitch angle and the yaw angle are measured by means of optical encoders. The pitch and yaw rates are obtained using numerical differentiation of the respective encoder readings. Letting , a simple mathematical model of the 2-DOF helicopter of the form (1) has been derived using Lagrange\u2019s equations of motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002124_j.triboint.2010.05.005-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002124_j.triboint.2010.05.005-Figure14-1.png", + "caption": "Fig. 14. Influence of loads on the film thickness, r\u00bc528 mm, ue\u00bc258 mm/s, PB1300.", + "texts": [ + " 10, the distance between the Hertzian contact center and the rotation axis of the glass is kept unchanged, so did the entrainment speed at the contact center. According to Eq. (1), with the increase of the contact radius with increase in loads, local entrainment speed at side-lobe R decreases and the side lobe film thickness hminR decreases as shown in Fig. 10. At side-lobe L, local entrainment speed increases with increase in load, but the measured hminL does not increase accordingly. The numerical results in Fig. 14 present similar film thickness variations with loads, correlated to the experiments in Fig. 10. Quantitatively, the numerical analyses give lower film thickness, but similar indices for film thickness at side lobes as the experiments are. EHL films under conditions of pure rolling with spinning have been obtained experimentally and numerically, and the effects of the spinning on the film generation are studied. The main results are summarized as follows: (1) In the present ball-on-disc configuration, different spin levels can be incorporated into the EHL under pure rolling by offset of the contact center with respect to the disc rotation center" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003105_j.ijsolstr.2011.09.017-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003105_j.ijsolstr.2011.09.017-Figure1-1.png", + "caption": "Fig. 1. Generalized Ziegler model subjected to partial follower load.", + "texts": [ + " In the next sections, the theoretical results derived in Section 2 are applied to a meaningful academic case (the generalized Ziegler column problem in Section 3), and then to a practical case (aircraft wings with aeroelastic effects in Section 4). An illustrating example is given with the nonconservative generalized Ziegler column, loaded by a partial follower load (Hermann and Bungay, 1964; Leipholz, 1987). The structure is composed of two bars AB and BC of length L, articulated on extremities A and B with a torque bending k. A force F is imposed at extremity C, with an angle ah2 with the vertical direction y. The rotation of each bar is described with angles h1 and h2 (Fig. 1). Henceforth, the range of a is restricted to [0,1]. Given a force F with an inclination ah2, the question of the existence of other geometrical configurations in the vicinity of the trivial solution h1 = 0 and h2 = 0, corresponding to the structure equilibrium, arises. This geometrical configuration is defined with angles h1 and h2. Assuming that both angles h1 and h2 are small with respect to 1, Eq. (9) yields: 2 p ap 1 1 1 \u00f01 a\u00dep h1 h2 \u00bc 0 0 \u00f038\u00de where p \u00bc FL k . K \u00bc 2 p ap 1 1 1 \u00f01 a\u00dep is the stiffness matrix" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001230_j.fusengdes.2008.11.021-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001230_j.fusengdes.2008.11.021-Figure2-1.png", + "caption": "Fig. 2. Exploded upside-down view of main UPL components.", + "texts": [ + " As the document is not yet available, he ITER-NL team adheres to the following list of self-imposed RH uidelines: Apply KISS (Keep It Simple, Stupid) [5]. Apply gravity stabilized handling of components. Minimize the number of tools and other equipment. Aim for communality of tooling between (remote) tasks where possible. - Apply captive fixings on any interface that is to be RH. - Provide alignment features on internal and external interfaces. 5. The RH concept for the UPL As a starting point for the RH concept design the UPL is assumed to be present in the hot cell. The main components to be maintained are identified in Fig. 2 (extracted from the UPL PDF). 6. VR applied to the RH concept Below a description in some detail is made as a concept for the replacement task of the blanket shield module (BSM) for the ECH-UPL. This task has strong commonalities with other Upper Port Plugs. The BSMs require maintenance (RH class 2). A pilot VR simulation for the BSM replacement is set up to initiate a first design iteration with strong RH involvement. The virtual environment model is based on the ITER 2005 Hot Cell, slightly modified by Oxford Technologies Ltd" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002632_j.mechmachtheory.2013.02.010-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002632_j.mechmachtheory.2013.02.010-Figure3-1.png", + "caption": "Fig. 3. Lumped parameter model of multiple mesh gear system driven by four pinions.", + "texts": [ + " A dimensionless dynamic model for an example gear system driven by two pinions is presented to investigate the correlation of the load sharing behavior and mechanical parameters such as the meshing frequencies, bearing stiffnesses, mounted locations and gear configuration coefficients. A photo of a multiple mesh gear system is shown in Fig. 1, in which the big external gear is actuated by the four surrounding pinions simultaneously. For every mesh gear pair, the meshing force is assumed as acting on the line of action. The corresponding lumped parameter model of the individual meshing gear pair can be set up as a spring\u2013damper system as illustrated in Fig. 2. Consequently, the lumped parametermodel for this gear systemdriven by four pinions can be summarized as shown in Fig. 3. Here, the pinions and the gear are fixed on the same platform. A synchronous control system is used for the pinions. Consequently, the difference of the mesh phases for the various meshing pairs is not considered in this context. The initial mesh phases for the gear pair are identical. The equations of motion with respect to the angular rotations of the n pinions and the gear, \u03b8pi and \u03b8g, can be written as Ipi\u20ac\u03b8pi \u00fe ciRpi Rpi _\u03b8pi\u2212Rg _\u03b8g \u00fe ki t\u00f0 \u00deRpi Rpi\u03b8pi\u2212Rg\u03b8g \u00bc Tpi; Ig\u20ac\u03b8g\u2212Rg\u2211ci Rpi _\u03b8pi\u2212Rg _\u03b8g \u2212Rg\u2211ki t\u00f0 \u00de Rpi\u03b8pi\u2212Rg\u03b8g \u00bc \u2212Tg ; \u00f01\u00de the i denotes the ith pinion and i = 1,\u22ef,n" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002887_1.4024547-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002887_1.4024547-Figure1-1.png", + "caption": "Fig. 1 Schematic view of three-pad foil bearings with preload; CS is the minimum radial clearance at the converging end of the film when rotor is at the center of the bearing", + "texts": [ + " Air/gas-lubricated bearings are environmentally friendly and allow compact system configurations with higher system efficiencies than similar systems with conventional bearings, because of much lower bearing friction losses. Air/gas bearings can run at extreme temperatures without the risk of thermal degradation of the lubricant, and load-carrying capacity increases with rotational speed, which is ideal for small turbomachinery operating at very high speeds. Foil bearings (FBs) using air or gas as a lubricant with compliant bearing surfaces have the added benefit of shock tolerance and better accommodation of minor shaft-bearing misalignments and shaft expansion due to thermal load. Figure 1 shows the schematic view of a FB with three top foils, corrugated bump foils, and the bearing sleeve. The bump foil provides stiffness and damping for the bearing, which can each be controlled by tailoring the bump height, materials, and the overall geometry. Typically, nickelbased superalloys are used for the foils, because of their excellent corrosion resistance and mechanical properties at high temperature. The inner surface profile of the entire top foil can be either circular or noncircular, depending on load-carrying capacity requirements and rotor-bearing stability. The three-pad FB shown in Fig. 1 consists of three top foils, with the top foil having an arc angle of 115 deg, supported by separate bump foils. Each top foil has a finite hydrodynamic preload (defined below). Because both the thin top foils and the bump foils are manufactured through cold-forming and heat treatment, the foils are always slightly loose from the sleeve, except for the fixed foil ends, regardless of the FB type. When the FB is inserted onto a shaft, the loose foil structure acts as a soft mechanical preload on the shaft without any visible clearance. However, if one assumes that the foil structure sits on the sleeve ideally, a finite assembly clearance exists. Once hydrodynamic pressure begins to develop, net bearing clearance is the sum of the assembly clearance and bump deflection. In Fig. 1, where the foil structure sits on the sleeve ideally, the center of each top foil is offset from the global bearing center by a small distance, rp, which is called the hydrodynamic preload. This arrangement gives a nonuniform assembly clearance around the circumference of the bearing, with a maximum radial clearance at the leading and trailing edges of the top foil and a minimum radial clearance, CS, at the center of the top foil. Then one can define the nondimensional hydrodynamic preload as Rp \u00bc 1 CS rp \u00fe CS (1) Friction and low load capacity during start/stop are generally not problematic if the static load on the bearing due to the rotor weight is not significant" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003907_iecon.2013.6699627-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003907_iecon.2013.6699627-Figure3-1.png", + "caption": "Fig. 3. Flux linkage calculation when if = 6A:(a) Mesh for FEM (b) Flux linkage of stator coils", + "texts": [ + " (4) where ved, veq are the d and q-axes voltage and \u03c9e is the electrical angular frequency, the superscript e represents a synchronous(rotating) reference frame. The torque equation is obtained as T = 3 4 P [ \u03bbe di e q \u2212 \u03bbe qi e d ] , (5) where P is the number of poles. A schematic diagram showing a high frequency signal injection to the field winding of an ESM is depicted in Fig. 1. The inductances can be decomposed as the sum of mutual and leakage parts: Ldd = Ldf + Lls (6) where Lls is the leakage inductance of stator. Using (3)- (4) and (6), an equivalent circuit can be depicted as shown in Fig. 2. Fig. 3(a) shows the mesh plot of the ESM used in this experiments, whose parameters are listed in Table I. Fig. 3(b) shows the changes of the flux linkage between a, b, c-phase windings and the field winding as the rotor rotates. Fig. 4 and Fig. 5 show the plots of \u03bbe d(i e d, i e q, if ) and \u03bbe q(i e d, i e q, if ) over the (ied, i e q) plane when if = 6A, which were obtained through two-dimensional(2D) FEM calculation. As shown in [6] [7], the incremental inductances are calculated according to Ldq \u2248 \u03bbe d(i e d, i e q +\u2206ieq, if )\u2212 \u03bbe d(i e d, i e q, if ) \u2206ieq Ldf \u2248 \u03bbe d(i e d, i e q, if +\u2206if )\u2212 \u03bbe d(i e d, i e q, if ) \u2206if " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001368_j.mechatronics.2010.09.010-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001368_j.mechatronics.2010.09.010-Figure2-1.png", + "caption": "Fig. 2. INU corridor for terrain elevation profile generation.", + "texts": [ + " This could be due to imperfect lateral track control of the vehicle, disturbances such as winds or gusts, and uncertainty or errors in the knowledge of the position of the vehicle. The positional (navigation) errors come about because of the inertial navigation unit (INU) onboard the vehicle. GPS aiding prevents the errors from growing too large but in times of GPS outage, the inertial system errors tend to diverge. Low cost INUs usually employed in UAVs have low accuracy inertial sensors and the rate of error growth with time is therefore high. Uncertainty in the knowledge of the position of the air vehicle thus generally increases with time. This is shown in Fig. 2 where the corridor of uncertainty centered around the nominal path of the vehicle is shown to increase with time (and distance traveled). Each circle indicates the uncertainty in the position of the vehicle at that range (or distance). The radii are continuously increasing since the error in the navigation solution grows with time (and range), and therefore the uncertainty in position also grows. Terrain profile is data containing terrain elevation versus range information. This information is used by the terrain-following algorithms. However due to the growing position uncertainty of the vehicle as indicated in Fig. 2, we need to find a robust elevation profile so that no peaks in the probable region of presence of the vehicle are missed out. Therefore for a given range, one should pick the highest value in the region so that chances of ground collision are ruled out. This is done as follows: The entire route is divided into a number of small segments and arcs. For straight part of the route, a length of say 100 m can be used. For turns a resolution of 100/Rt radians (equivalent to 100 m) can be used, Rt being the turn radius of the vehicle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure7-1.png", + "caption": "Fig. 7 Only two CRCMs are needed to balance the moment, and the other two remain as CMs \u202014\u2021", + "texts": [ + " This means that there are only two CRCMs necessary, which oth can be constructed compactly near the base, as shown in Fig. . This is a configuration described in Ref. 14 . The former RCMs become fixed countermasses. It is also possible to derive this parallel mechanism by combiation of an idler loop 15 and a CRCM-balanced double penduum, as shown in Fig. 8. Also in this case, only two CRCMs are ecessary, and by using the countermass of the idler loop as a RCM, they can be constructed near the base. This result is hown in Fig. 9 and it has only one fixed countermass instead of wo, as in the configuration of Fig. 7. With the equations of the angular momentum of the balanced ouble pendulum being known, the inertias of the CRCMs and the nertia equations of the mechanism can be calculated quickly by imply adding the equations of each individual double pendulum. s an example, the angular momentum and inertia equations of A O ig. 6 2DOF balanced parallel mechanism obtained by combiation of two CRCM-balanced double pendula he parallel mechanism of Fig. 7 are calculated. Therefore, it is ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201 assumed that half of the mass m2 and inertia I2 at the end point is balanced by each double pendulum. By changing the notations slightly with m2 =m2,1 , m1 =m1,1 , I2 = I2,1 , and I1 = I1,1 in which the additional index 1 represents double pendulum 1, the angular mo- NOVEMBER 2009, Vol. 131 / 111003-5 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use i C r A r f 4 u f t d d b C c s t a n fi w t t 1 Downloaded Fr + I2 2 + m2 2 l1 2 + m1,2 l1 2 + I1,2 \u03071 44 n which links 1 and 2 have lengths l2 and l1, respectively", + " 50 and 51 then esults in hO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a + I2 + m2l2 2 + m2 l2 2 + k2I1,b \u2212 l1 cos 1 l2 cos 2 \u2212 1 \u03071 53 he single equation of the reduced inertia is now dependent on the osition of the mechanism. This equation also holds for an uncontrained balanced double pendulum moving along the same trajecory, although then there are two input angles where each has a onstant reduced inertia. Another balanced slider mechanism can be derived from the arallel mechanism of Fig. 7. If the end point of this parallel echanism moves along a straight line through the origin, as hown in Fig. 13, then the two CRCMs can become fixed counermasses. Half of the mass m2 and half of I2 if not a slider is hen balanced by each link that is attached to it. The inertia and ength of these links must be equal and also the links attached to he origin must have equal inertia and length by which they beome moment balanced sets. With these conditions, it is also posible to have the length and inertia of the links attached to the lider to be different from the length and inertia of the links atached to the origin" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001654_10426914.2010.489593-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001654_10426914.2010.489593-Figure1-1.png", + "caption": "Figure 1.\u2014Schematic diagram of metal deposition in a narrow weld groove with the formation of cavity in weld pool.", + "texts": [ + " The studies have been carried out using 25mm thick control rolled micro-alloyed grade SAILMA-410HI/SA533 high-strength low-alloy (HSLA) steel plate considering its large attraction to industries in construction of various advanced welded structures. Spatial model for estimation of size and thermal behavior of weld pool in single seam per layer multipass narrow gap welding In order to produce a weld joint free from lack of groove wall fusion by using single seam per layer multipass narrow gap welding, the weld pool temperature should be able to fuse it apart from melting a part of weld deposit of earlier pass. It may be achieved by having a weld pool of required geometry and temperature primarily depending upon the design of weld groove (Fig. 1) accomodating and allowing the weld metal to come in contact with the groove wall. The critical geometry and the temperature of the weld pool necessary to produce single seam per layer multipass narrow groove weld free from lack of groove wall fusion can be estimated as follows. The geometrical aspects of weld pool in single seam per layer multipass narrow gap welding primarily concern its ability to intimately contact the groove wall in order to result its desired fusion and filling the fraction of weld groove as estimated by their proportionate cross-sectional area", + " (17) and (18) for the square of ratio of R1 and R2, by considering d/h << 1, may be written as( R1 R2 )2 = 1( 2\u00b7h a1 )2 + ( b1 a1 )2 (19) As the values of both the ratios of h/a1 and b1/a1 are less than one, the values of R1/R2 will be greater than one, which shows that R1 shall always be higher than R2. Thus, the point P1 at a distance R1 from the arc center shall predominantly play the critical role to assure required fusion of the groove wall. The heating of weld pool under P-GMAW is primarily attributed to arc heat and heat transfer by the superheated filler metal under pulsed current. The arc heat acting continuously on the system as double ellipsoidal heat source [12] is defined by three-dimensional ellipsoidal heat source parameters (Fig. 4). The arc may cover a part of the weld pool (Fig. 1) from where the heat is transferred to rest of the weld pool primarily by conduction and convection of liquid metal. The arc heat transferred to the point P1 of the weld pool is contributing to raise the temperature of groove wall from initial preheating temperature upto a temperature of Tarc, which may cause an initial fusion to it. The heat transferred by the superheated filler metal as point heat source, acting at the depth of cavity formed on the weld pool contributes additional heat to the point P1 to raise its temperature further to Tfiller", + " D ow nl oa de d by [ L au re nt ia n U ni ve rs ity ] at 0 2: 41 0 8 O ct ob er 2 01 3 Figure 18.\u2014Comparison of the estimated and required temperature of the weld pool at different Im and , respectively of (a) 220A and 0.25, (b) 240A and 0.25, and (c) 240A and 0.15 at of 10 2\u00b1 0 4 kJ/cm. But, at a relatively higher 13 4\u00b1 0 5 kJ/cm, it is observed Figs. 19(a)\u2013(d) that the required weld pool temperature is lower than the estimated one even up to six weld passes in the narrow groove, as shown in Fig. 1. Thus, here it may be expected that the narrow groove as shown in Fig. 1 can be filled up without any lack of fusion by single seam per layer multipass weld deposition. In order to validate the anticipation of weld quality as a result of the observations made in Figs. 18 and 19, some multipass narrow groove (Fig. 1) P-GMA welding of 25mm thick control rolled micro-alloyed HSLA steel plate of grade SAILMA-410HI/SA533 was carried out by multipass centrally laid single seam per layer vertically down weld Figure 19.\u2014Comparison of the estimated and required temperature of the weld pool at different Im and , respectively of (a) 265A and 0.08, (b) 265A and 0.15, (c) 265A and 0.25, and (d) 240A and 0.25 at of 13 4\u00b1 0 5kJ/cm. D ow nl oa de d by [ L au re nt ia n U ni ve rs ity ] at 0 2: 41 0 8 O ct ob er 2 01 3 Figure 20" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000883_s11661-008-9637-8-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000883_s11661-008-9637-8-Figure2-1.png", + "caption": "Fig. 2\u2014Schematic diagram of bead-on-plate weld deposit.", + "texts": [ + " The geometry of the weld deposit, defined by the depth, width, and area of the fusion of the base plate, the bead height (reinforcement), the bead width, the weld toe angle, and the area of the weld deposit, was measured on the computer-scanned image of the metallographically prepared and etched specimens, with the help of Adobe Photoshop 7.0 software (Adobe Systems Incorporated, San Jose, CA). The accuracy of the measurements was also confirmed by studying the same factors under an optical microscope at a suitable magnification. The toe angle of the bead deposit was measured in reference to a tangent to the curvature of the bead reinforcement drawn at a point 1 mm above the plate surface, as is schematically shown in Figure 2. The fusion area of the base plate and the area of the weld deposit were measured through graphical methods, for more accuracy, especially in the case of the relatively nonuniform fusion geometry. The form factor (FF) of the weld deposit was estimated as FF \u00bcWb H \u00bd13 where Wb is the bead width and H is the bead height (reinforcement). Assuming that the amount of the weld deposition and base metal melting are a linear METALLURGICAL AND MATERIALS TRANSACTIONS A VOLUME 39A, DECEMBER 2008\u20143263 function[19] of the area of the weld deposit and the area of fusion, respectively, in the bead-on-plate weld deposition, as is schematically shown in Figure 2, they were estimated on the transverse section of the weld deposit. The fraction (F) of base metal fusion per unit mass of the weld deposition and the dilution (D) were estimated as F \u00bc Afvqb Advqd \u00bd14 D \u00bc Af Af \u00fe Ad \u00bd15 where Af is the area of fusion of the base plate, Ad is the area of the weld deposit, and qb and qd are the density of the base and deposited metals, respectively. The reported value of each aspect of the measured geometry of the welds prepared at different parameters is the mean of a number of welds prepared at each parameter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000341_iecon.2007.4460323-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000341_iecon.2007.4460323-Figure3-1.png", + "caption": "Fig. 3. Flux linkages with respect at air gap of a PMSM 50 % magnetized.", + "texts": [ + " 6 show the stator cunrent harmonic obtained from simulations of a PMSM with demagnetization. It is shown that the amplitudes of harmonics first and fifth are higher in a faulty machine than in a healthy PMSM. However, this criterion is difflcult to apply for a low speed. Any possible asymmetry induces currents in windings or stray paths and variations of magnetic properties; on the other hand, tolerances in physical dimensions may cause unequal flows of fluxes on the left-hand and the right-hand sides, as it is shown in Fig. 3. The distributed MMF is not sinusoidal in case of demagnetized machine. Thus. the amplitude of MMF falls to constant amplitude lower than the nominal in the pole pair below fault. This variation can be considered as a pulse width modulation with a frequencyf, and duty cycle 1/p, as it can be seen in Fig. 4. The variable demagnetization MMF is calculated assuming the term ofpulse width modulation has the form. F, = Aod sin(p09\u00b1w3(1\u00b1k/p)t) (1) Thus, the MMF under failure together with the constant permeance induces currents at frequencies ofmain multiples. Now, if we consider that the MMF interacts with certain dynamic eccentricity, due to either the demagnetization or change of permeance for the pair of poles under failure, the flux density is represented by the following expression (see Fig. 3). B:sd Adsin (PC TI)O\u00b10l\u00b1+ k\u00b11 P.? Jt (2) However, if demagnetization exists, low-frequency components near the fundamental appear [9], given by: fd11g = fA \u00b1 k p k = 1,2,3,... (3) In case of constant speed at high and medium ranges, FFT allows detecting demagnetization fault analyzing the amplitude of harmonics 1st and 5th only, especially for high rotor speed. However, fault detection by FFT is not clear at low speed. Furthermore, it is not possible to apply FFT in case of speed and torque variations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002525_tac.2012.2186183-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002525_tac.2012.2186183-Figure4-1.png", + "caption": "Fig. 4. Gaze motion from (25) for increasing values of the parameter , with the external control set to zero, i.e., , and the motion is purely due to the potential function (21), which has a minimum at the frontal gaze. (a) . (b) . (c) . (d) .", + "texts": [ + " In Section V, we define a Lagrangian using the Riemannian Metric in Section IV and a suitable Potential Energy term, and write down the corresponding Euler Lagrange Equations (EL). We also add an external \u201cgeneralized torque\u201d as inputs to the EL equations and this way we obtain a controlled dynamical system. In Section VI, we illustrate the obtained geodesic equations and the controlled dynamical system equations by a set of simulations. In these simulations, the control is either set to zero (see Fig. 4) or chosen to simulate an appropriate damping (see Figs. 5 and 6). The control system is regulated by an appropriate choice of a Potential Energy term. Finally Section VII concludes the paper. In Appendix I, we state and prove the Donders\u2019 Theorem for head movement which generalizes Listing\u2019s Theorem (see Section III) for eye movement. In Appendix II, we state and prove the half angle rule under Donders\u2019 constraint which generalizes the half angle rule for Listing\u2019s constraint. The half angle rule describes constraints on the angular velocity vectors and angular acceleration vectors, when the motion dynamics satisfy Listing\u2019s and Donders\u2019 laws", + " Our next goal is to show that by tuning the parameters and in (21), we can drive the eye or the head to a suitable end position and orientation. This has been illustrated in the next three examples. The purpose of this subsection is to demonstrate via simulation that by adding a potential term, one is able to push the trajectories of the eye or head toward the frontal gaze direction. Example 2.1 (Eye Motion With a Potential Function But No Damping): In this example we solve (25) on and display the gaze trajectories in Fig. 4. We have assumed , and in (21). Increasing magnitudes of has been chosen in Fig. 4(a)\u2013(d). Our simulations show that with increasing magnitude of the potential function, the gaze trajectories are restricted to a smaller neighborhood of the frontal gaze. However, the trajectories are oscillatory. We now proceed to add a damping term to the motion equations. Example 2.2 (Eye Regulation Toward Frontal Gaze Direction Using a Potential Function and Damping): We repeat Example 2.1 but choose and in (25), in order to dampen the fluctuations in the trajectories. The results are plotted in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003637_jahs.57.012004-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003637_jahs.57.012004-Figure3-1.png", + "caption": "Fig. 3. Airfoil deflection test illustration.", + "texts": [ + " The three suspending points and CG are in the same plane, which is perpendicular to the selected body-frame axis. HeLion is then swung around the body-frame axis, and the recorded torsional oscillation period is used to compute the moment of inertia (category B in Table 3). Airfoil deflection test. This test aims at determining the parameters related to the Bell\u2013Hiller mixer. For both longitudinal and lateral directions, three experiments (similar to Ref. 23) have been conducted. Taking the longitudinal direction as an example, the experiment procedure is depicted in Fig. 3. Three parameters, i.e., Alon, Clon, and Ksb, can be determined. 1) Determination of Alon (see Fig. 3(a)) consists of three steps: (a) adjust and maintain the stabilizer bar to be level to the X axis of the body frame, (b) inject \u03b4lon to tilt the swash plate longitudinally, and (c) record \u03b8cyc,as (the cyclic pitch deflection of the main rotor blade). Alon is the ratio of \u03b8cyc,as to \u03b4lon. 2) For Clon (see Fig. 3(b)), the experiment follows: (a) adjust the stabilizer bar to be level to the X axis of the body frame, (b) keep the cyclic pitch of the main rotor blade unchanged, (c) inject \u03b4lon to tilt the swash plate longitudinally, and (d) record the deflection of the stabilizer bar paddle. Clon is the ratio of the paddle deflection to \u03b4lon. 3) For Ksb (see Fig. 3(c)), the experiment consists of (a) adjust the stabilizer bar to be level to the X axis of the body frame, (b) keep the swash plate balanced, and (c) manually change the stabilizer bar flapping angle cs and record the corresponding change in \u03b8cyc,as . Ksb is the ratio of \u03b8cyc,as to cs. The same experiments are applied to the lateral direction to identify Blat and Dlat. Another numerical result of Ksb can be obtained via a lateral deflection experiment. The two results are almost identical, which is consistent with the expectation that Ksb is a Bell\u2013Hiller mixer setting that is strictly symmetrical to both directions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000726_icems.2009.5382754-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000726_icems.2009.5382754-Figure5-1.png", + "caption": "Fig. 5. Analytical model of segment type SRM.", + "texts": [ + " mqqN \u2264+ (5) These equations show that \u03b1 can be shorten rather than \u03c6 by increasing the N greater than 3 and the m greater than 4. If two or more phase windings are excited together at start and end of excitation, the torque can be superimposed and so the torque ripple will be reduced. In this paper, we analyze the torque characteristics when the number of phases is 4, 5 and 6. Table 1 shows \u03c6, \u03b4, and \u03b1 above conditions. The \u03b1 becomes smaller than the \u03c6 when the number of phases is 4, 5, and 6, and the two or more phase windings can be excited at the same time. IV. ANALYTICAL MODEL Figure 5 shows the analytical models of a 4-pase machine, a 5-pase machine and a 6-phase machine. Diameter of the stator is 160mm, diameter of the rotor is 90mm, gap length is 0.3mm and they are the same as the 3-phase machine. Figure 6 shows the exciting current waveforms of them. The excitation conditions are shown in Table 1. The number of turns of the stator windings is 75 turns in the 4-phase machine, 90 turns in the 5-phase machine and 113 turns in the 6-phase machine so that a total magneto motive force equals to that of the 3-phase machine" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure5-1.png", + "caption": "Figure 5: Twin Toroidal Transmission", + "texts": [ + " The first Series Production of the Torotrak full toroidal Variator technology is as part of a transmission system for a ride on lawntractor in the Outdoor Power Equipment (OPE) market. To compete in this particularly cost sensitive market segment, significant simplification was required in both the design and operation of the Variator. This simplification was achieved by utilizing a single cavity design, reducing the number of rollers in the cavity from three to two and introducing a simple \u2018yoke\u2019 style roller control mechanism. This two roller design is in Series Production in the Twin Toroidal Transmission (TTT) as shown in Figure 5. With the two roller technology now validated, a simple, low-cost variable drive system has been created. One application of such a variable drive is optimization of a supercharger. The increasing demand for improving vehicle fuel economy is guiding automakers into developing more efficient powertrain systems. One area receiving significant attention is the optimisation of the base engine itself where both downsizing and reducing the pumping losses of the engine are essential. Therefore, pressure charging the engine, either via a supercharger or a turbocharger, is becoming increasingly prevalent in both today\u2019s and in particular future automobiles" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000871_09544062jmes949-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000871_09544062jmes949-Figure7-1.png", + "caption": "Fig. 7 FEM of the FS", + "texts": [ + " The FEM takes these effects into account during the simulations. The software used in this research allows the use of axial elements that can be programmed to characterize a non-linear stiffness. However, axial elements with linear stiffnesses are used in the current FEM to model the balls to reduce the computational time, given that a high number of contact elements are used and that these elements already required a non-linear solution path. The teeth of the FS are modelled with eight-node brick elements (seedetails inFig. 7). Contact elements, placed on the external surfaces of these bricks, make it possible to control the sliding between the teeth of the FS and those of the CS. The remaining portion of the FS is modelled with shell elements as these elements are best suited for thin sections. Constraint equations between the solid elements of the teeth and the shell elements of the cup are used in order to allow the parts to remainmated under deformation of the FS. Numerical experiments have shown that FEM involving contact surfaces is time-consuming because the associated analyses are highly non-linear [14]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000924_978-3-540-30301-5_57-Figure56.35-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000924_978-3-540-30301-5_57-Figure56.35-1.png", + "caption": "Fig. 56.35a\u2013d Climbing a ladder. (a) The robot begins to climb; it has contacts on both hands and feet. (b) The right foot is then controlled to move up one step. (c) Next, the center of mass is controlled to move to the right in order to maintain balance with two hands and the right foot. (d) The left foot is then controlled to move up one step [56.84]", + "texts": [ + " To achieve this control hierarchy, operational tasks are projected into the constraint null space, and postures are projected into the task null space followed by the constraint null space (Fig. 56.34). Unlike resolved momentum control, which generates joint velocities, all of these controllers generate joint torques. When calculating these torques, the controllers use models of the dynamics of the robot and the environment. A ladder-climbing behavior tested in simulation demonstrates this framework for control (Fig. 56.35). In this example, the desired trajectories for the center of mass, the hands, and the feet are specified in advance. When the simulation is run, the whole-body control system generates joint torques that seek to meet these coarse motion specifications, which results in the simulated robot climbing the ladder. While climbing, the simulated robot successfully resists unexpected disturbances. Research into methods for the coordination of controllers for humanoid robots is an active area of research, for example, Mansard et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001100_j.fss.2009.12.004-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001100_j.fss.2009.12.004-Figure1-1.png", + "caption": "Fig. 1. Satellite reference and body coordinates.", + "texts": [ + " An SDRE and SMC controllers are described in Section 5 and Appendix B, respectively, followed by simulations and comparison of controllers\u2019 performance in Section 6. Three degree of freedom (3-dof) rigid satellite model is presented in this section. The attitude motions about the three principal axes are nearly uncoupled, and stabilization about each principal axis may be treated separately. Axes X B , YB , and Z B define the satellite\u2019s body axis frame, and the axis system is considered centered at the center of gravity as shown in Fig. 1. Thrusters are available to produce torques about each of the three principal axes. The physical interpretation of the Euler angles [18] for a micro-satellite platform is illustrated in Fig. 1. The roll ( ), pitch ( ), and yaw ( ) angles are defined by successive rotations around the coordinate axes X B, YB, and Z B in the body fixed frame. For large angles and position control [7,4] of spacecraft, quaternion rotation is used. This is more suitable as it avoids singularity functions and gimbal lock [9]. Euler angular moment (H\u0307 ) [18] which performs satellite attitude-rotational motion in space is expressed in the form of a matrix as follows: M = d H dt + [ \u00d7 H ] d H dt = M \u2212 [ \u00d7 H ] because H = I \u00b7 I \u00b7 d dt = M \u2212 [ \u00d7 I \u00b7 ] d dt = \u23a1 \u23a2\u23a2\u23a3 \u00b7 p \u00b7 q \u00b7 r \u23a4 \u23a5\u23a5\u23a6 = I \u22121 \u00b7 M \u2212 [I \u22121 \u00b7 ] \u00d7 [I \u00b7 ] (1) where vector M , the applied moment (thrusters), is the input u, vector is the angular rate and matrix I is the inertia around the body principal axes X B , YB , and Z B " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003110_s12239-012-0022-7-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003110_s12239-012-0022-7-Figure7-1.png", + "caption": "Figure 7. Numerical model for the journal bearing between the connecting rod and crankshaft.", + "texts": [ + " The EHD and MFBD solvers are used iteratively to implement these procedures to analyze the lubrication and dynamic characteristics of the journal bearing. To support a general-purpose EHD solution, groove and oil hole effects are implemented as pressure boundary conditions in the EHD solver. To implement the EHD module with the MFBD solver, this study used the RecurDyn\u2122 (2010) MFBD environment. To validate the numerical results of this study, the experimental results of Nakayama et al. (2003) were used. The numerical model has been described in detail in Nakayama et al. (2003). Figure 7 shows the numerical model and the measurement points of the oil film thickness. To measure the effect of the external load in the model, loads of 100 N, 200 N, 500 N, 1000 N, 1500 N, 2000 N, and 2500 N were applied. The rotational speed of the crankshaft was 3570 rpm. Table 1 shows the simulation parameters used in the numerical model. Figure 8 shows the results for the pressure distribution according to the rotational angle for the grid points on the center circle. The pressure peak increases with increasing force, and the region of raised pressure is from about 50\u2013180 deg, as expected" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure11-1.png", + "caption": "Fig. 11 Balanced crank-slider mechanism synthesized from the CRCM configuration with a single CRCM", + "texts": [], + "surrounding_texts": [ + "a o d d f t b\nr o l m\n3\nm p t q\np p o n n p b n\nm s b 7 C\nn l n C s t\nd i s A t\nF n\nJ\nDownloaded Fr\nparallel transmission k1=0 and does not rotate for any motion f the mechanism. m2\nis then used for the moment balance of both egrees of freedom, as in Fig. 4. The result is that the term I1 rops from Eqs. 26 , 28 , 29 , 31 , and 32 . It is evident that or a low inertia, the countermasses should not rotate with respect o the base, besides what is necessary to maintain the moment alance.\nLimitations of the CRCM principle can be the transmission atios that have to remain relatively small in practice, a gear ratio f 8 is already high , and the shape of a CRCM that may become arge to obtain the specific amount of inertia and mass. These and\nore limitations are studied in Ref. 12 .\nCRCM-Balanced 2DOF Parallel Mechanisms In this section, three new CRCM-balanced 2DOF parallel echanisms are synthesized by using the CRCM-balanced double\nendulum of Fig. 1. It is shown that the balancing conditions and he inertia equations for these parallel mechanisms can be derived uickly.\nAny combination of one or more balanced double or single endula results into a balanced mechanism. Therefore, two double endula balanced, as in Fig. 1, can be combined, such that their rigin is at the same location and they form the parallel mechaism of Fig. 6. The end point of each double pendulum does not eed to coincide but can be anywhere as long as the links remain arallel. The mass at the end point and its inertia can be balanced y both links or by only one of them. The other then is still ecessary to balance the mass and inertia of the link itself.\nSince the angular velocities of parallel links are equal for the oment balance of two parallel links, only one CRCM is necesary. This means that there are only two CRCMs necessary, which oth can be constructed compactly near the base, as shown in Fig. . This is a configuration described in Ref. 14 . The former RCMs become fixed countermasses. It is also possible to derive this parallel mechanism by combiation of an idler loop 15 and a CRCM-balanced double penduum, as shown in Fig. 8. Also in this case, only two CRCMs are ecessary, and by using the countermass of the idler loop as a RCM, they can be constructed near the base. This result is hown in Fig. 9 and it has only one fixed countermass instead of wo, as in the configuration of Fig. 7.\nWith the equations of the angular momentum of the balanced ouble pendulum being known, the inertias of the CRCMs and the nertia equations of the mechanism can be calculated quickly by imply adding the equations of each individual double pendulum. s an example, the angular momentum and inertia equations of\nA\nO\nig. 6 2DOF balanced parallel mechanism obtained by combiation of two CRCM-balanced double pendula\nhe parallel mechanism of Fig. 7 are calculated. Therefore, it is\nournal of Mechanical Design\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201\nassumed that half of the mass m2 and inertia I2 at the end point is balanced by each double pendulum. By changing the notations slightly with m2 =m2,1 , m1 =m1,1 , I2 = I2,1 , and I1 = I1,1 in which the additional index 1 represents double pendulum 1, the angular mo-\nNOVEMBER 2009, Vol. 131 / 111003-5\n6 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "i\nC r\nA r f\n4\nu f t d\nd b C c s t a n\nfi w t t\n1\nDownloaded Fr\n+ I2\n2 +\nm2\n2 l1 2 + m1,2 l1 2 + I1,2 \u03071 44\nn which links 1 and 2 have lengths l2 and l1, respectively. \u03071 and\n\u02d9 2 depend on \u03071 and \u03072 as\n\u03071 = \u2212 \u03072 45\n\u03072 = \u03071 + \u03072 46\nombining 1hO,z and 2hO,z and substituting Eqs. 45 and 46 esult into one equation for the angular momentum:\n1+2 hO,z = I2 + I2,1 + I1,2 + m2 + m1,2 l2 2 + m2,1 + m2,2 l2 2\n+ m2 + m2,1 l1 2 + m1,1 + m1,2 l1 2 + k1I1,1 + k2I2,2 \u03071\nI2 2 + I2,1 + m2 + m1,2 l2 2 + m2,1 + m2,2 l2 2 + k2I2,2 \u03072 47\ns before, from the angular momentum, the equations for the educed inertias I 1 red and I 2\nred can be derived. The conditions for the orce balance of the mechanism in this case are\nm1,1 = m2\n2 + m2,1\nl1 l1 , m2,1 = m2l2 2l2\nm1,2 =\nm2l1\n2l1\n, m2,2 = m2\n2 + m2,2\nl2\nl2\nCrank-Slider and Four-Bar Mechanisms In this section, the various CRCM-balanced double pendula are sed to derive CRCM-balanced crank-slider mechanisms and our-bar mechanisms, and it is shown that for these mechanisms, he balancing conditions and the inertia equations can also be erived quickly.\nBy restricting the motion of the end point of the balanced ouble pendulum to move along a specific trajectory, CRCMalanced crank-slider mechanisms can be obtained from the RCM-balanced double pendula of Figs. 1 and 3\u20135. For the three onfigurations, the CRCM-balanced crank-slider mechanisms are hown in Figs. 10\u201312, for which the slider moves along a straight rajectory with offset h. The slider mass then does not rotate, and CRCM is only needed for the moment balance of the link conected to the slider. The advantages of each CRCM configuration remain if the con-\nguration is used as crank-slider mechanisms. Important features, ith respect to many other possible balancing configurations, are\nhat there are no transmission irregularities or singularities, and he mechanism can fully rotate by suitable link lengths .\n11003-6 / Vol. 131, NOVEMBER 2009\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201\nAlso for the crank-slider mechanisms, the conditions for the force and moment balance and the inertia equations can be obtained easily. The procedure is equal to that of the parallel mechanisms: first writing the angular momentum of the double pendulum, which is known, and then substituting the kinematic relations. In these 1DOF crank-slider mechanisms, 2 depends on 1. This relation is easy to find from the second equation of r2 and its derivative:\nr2,y = l1 sin 1 + l2 sin 2 = h 48\nTransactions of the ASME\n6 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "W w\nA a\nS r\nT p s t c\np m s t t l t c s s t\nC a\nF t p\nJ\nDownloaded Fr\nr\u03072,y = l1\u03071 cos 1 + l2\u03072 cos 2 = 0 49\nith \u03072= \u03071+ \u03072, \u03072 can then be written as\n\u03072 = \u2212 l1 cos 1\nl2 cos 2 \u2212 1 \u03071 50\nith\n2 = sin\u22121 h \u2212 l1 sin 1\nl2 51\ns an example, the configuration of Fig. 12 is taken for which the ngular momentum writes from Eq. 37 :\nhO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a \u03071\n+ I2 + m2l2 2 + m2 l2 2 + k2I1,b \u03072 52\nubstituting the kinematic relations of Eqs. 50 and 51 then esults in\nhO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a\n+ I2 + m2l2 2 + m2 l2 2 + k2I1,b \u2212 l1 cos 1\nl2 cos 2 \u2212 1 \u03071 53\nhe single equation of the reduced inertia is now dependent on the osition of the mechanism. This equation also holds for an uncontrained balanced double pendulum moving along the same trajecory, although then there are two input angles where each has a onstant reduced inertia.\nAnother balanced slider mechanism can be derived from the arallel mechanism of Fig. 7. If the end point of this parallel echanism moves along a straight line through the origin, as hown in Fig. 13, then the two CRCMs can become fixed counermasses. Half of the mass m2 and half of I2 if not a slider is hen balanced by each link that is attached to it. The inertia and ength of these links must be equal and also the links attached to he origin must have equal inertia and length by which they beome moment balanced sets. With these conditions, it is also posible to have the length and inertia of the links attached to the lider to be different from the length and inertia of the links atached to the origin.\nBerestov 10 showed a planar four-bar mechanism balanced by RCMs driven by inner gears. This mechanism can be regarded\nA\nO\nl *\n2\nl *\n1\nl 1\nl *\n1\nl 2\nl *\n2\nm 2\nl 1\nl 2\nig. 13 1DOF crank-slider mechanism without CRCMs obained by restricting the motion of the end point of the 2DOF arallel manipulator\ns a combination of a balanced single and a balanced double pen-\nournal of Mechanical Design\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201\ndulum, shown with chain driven CRCMs in Fig. 14. Also for four-bar mechanisms, the different CRCM configurations are applicable just as the substitution of the well-known kinematic relations into the inertia equations of the double and single pendula to obtain the inertia about one of the links. This means that with the equations for the double and single pendula and the kinematic relations the inertia of any four-bar mechanism can be written down easily.\nIt is a special case for which the four-bar linkage becomes a parallelogram. From Fig. 9, and assuming the link between O and A to be fixed with the base, the resulting parallelogram can be balanced, as in Fig. 15, with solely a CRCM. Because the coupler link does not rotate, its center of mass can be located arbitrarily.\n5 3DOF Parallel Mechanisms CRCM-balanced 3DOF planar and spatial mechanisms can be synthesized by combining the CRCM-balanced double pendula. Two examples are the planar 3-RRR parallel mechanism of Fig. 16, which has one rotational and two translational DOFs, and the spatial 3-RRR parallel mechanism of Fig. 17, which has two rotational and one translational DOFs. As described in Ref. 16 , the platforms of these mechanisms can be modeled by lumped masses at their joints, maintaining its original mass, its location of the center of mass, and its inertia tensor. This allows each leg to be balanced individually for which their combination is balanced too. The dimensions of each leg can be different, as long as each leg is balanced.\nTo obtain the inertia equations of these mechanisms, the kinematic relations can be substituted into the angular momentum equations of the double pendula. Since there are multiple closed\nNOVEMBER 2009, Vol. 131 / 111003-7\n6 Terms of Use: http://www.asme.org/about-asme/terms-of-use" + ] + }, + { + "image_filename": "designv11_3_0000558_jf902816e-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000558_jf902816e-Figure5-1.png", + "caption": "Figure 5. Interaction of different monoacid TAGs with reaction time on incorporation (mol %) of oleic acid. The reaction conditions are as follows: substrate mole ratio, 1:1; reaction temperature, 50 C; and enzyme dosage, 10%.", + "texts": [ + " Thus, the solid state could cause obstacles to the access of substrate to the active site of the lipase during acidolysis, even if it is solved in an organic solvent. Hence, higher oleic acid incorporation was observed when the binary mixtures of oleic acid and T6 or T8 were used. Reaction Time. Effects of various reaction times ranging from 3 to 24 h on the amount of oleic acid incorporated into different TAG structureswithLipozymeRMIMwere investigated. Threedimensional plots for the interaction of differentmonoacidTAGs with reaction time on incorporation (mol %) of oleic acid are shown in Figure 5. As the chain length of FA in TAG increased, except tristearin, the amount of oleic acid incorporated in TAGs varied from tricaproin to tristearin decreased at the longer reaction times. The amount of oleic acid incorporated into the same TAG was found to be statistically significant (P < 0.05) when the reaction time was increased. The highest increases were observed when the reaction time increased from 3 to 6 h for all TAG species. The reaction reached a steady state at 12 h of reaction for the binary mixtures of oleic acid and T16" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000253_978-1-4020-8600-7_33-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000253_978-1-4020-8600-7_33-Figure1-1.png", + "caption": "Fig. 1 3-RRR PPM with variable actuation.", + "texts": [ + " Rakotomanga et al. This section deals with the kinematic modeling of a new variable actuated mechanism (VAM), its singularity analysis, the presentation of some performance indices and the concept of regular dextrous workspace. The concept of VAM was introduced in [2, 15]. Indeed, they derived a VAM from the architecture of the 3-RPR planar parallel manipulator (PPM) by actuating either the first revolute joint or the prismatic joint of its limbs. This paper deals with the study of a new VAM illustrated in Figure 1. This mechanism is derived from the architecture of the 3-RRR PPM. As a matter of fact, the first link of each limb of the conventional 3-RRR manipulator is replaced by parallelogram AiBiDiEi to come up with the mechanism at hand. Accordingly, links AiBi and BiCi can be driven independently, i.e., angles \u03b1i and \u03b4i are actuated and uncoupled, by means of an actuator and a double clutch mounted to the base and located in point Ai , i = 1, 2, 3. 312 Kinetostatic Performance of a Planar Parallel Mechanism with Variable Actuation It turns out that the VAM has eight actuating modes as shown in Table 1", + " The unit is not specified as absolute lengths are not necessary to convey the idea. The velocity p\u0307 of point P can be obtained in three different forms, depending on which leg is traversed, namely, p\u0307 = \u03b1\u03071E(c1 \u2212 a1)+ \u03b4\u03071E(c1 \u2212 b1)+ \u03c6\u0307E(p\u2212 c1) (1) p\u0307 = \u03b1\u03072E(c2 \u2212 a2)+ \u03b4\u03072E(c2 \u2212 b2)+ \u03c6\u0307E(p\u2212 c2) (2) p\u0307 = \u03b1\u03073E(c3 \u2212 a3)+ \u03b4\u03073E(c3 \u2212 b3)+ \u03c6\u0307E(p\u2212 c3) (3) with matrix E defined as E = [ 0 \u22121 1 0 ] 313 N. Rakotomanga et al. ai , bi and ci are the position vectors of pointsAi , Bi and Ci , respectively. \u03b1\u0307i , \u03b4\u0307i and \u03c6\u0307 are the rates of angles \u03b1i , \u03b4i and \u03c6 depicted in Figure 1, i = 1, 2, 3. The kinematic model of the VAM under study can be obtained from Eqs. (1)-(c) by eliminating the idle joint rates. However, the latter depend on the actuating mode of the mechanism. For instance, \u03b4\u03071, \u03b4\u03072 and \u03b4\u03073 are idle with the first actuating mode and the corresponding kinematic model is obtained by dot-multiplying Eqs. (1)-(c) with (ci \u2212 bi )T , i = 1, 2, 3. Likewise, \u03b4\u03071, \u03b4\u03072 and \u03b1\u03073 are idle with the second actuating mode and the corresponding kinematic model is obtained by dot-multiplying Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001117_045104-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001117_045104-Figure3-1.png", + "caption": "Figure 3. Evaluation of uncertainty of gear measurement using the VGC.", + "texts": [ + " In the case of evaluating the uncertainty from the actual measurement result, the distribution of measured values is obtained from a number of measurement results. The extended uncertainty U is expressed as follows [23], where the coverage factor is k, standard uncertainty (the variability of a series of measurements) is um, reference artefact calibration uncertainty is un, geometry similarity influence is ug, work piece characteristic influence is uw and the bias of a series of measurement results is E: U = k \u221a u2 m + u2 n + u2 g + u2 w + |E|. (1) Figure 3 shows the concept of modeling the condition of the error factors of the actual gear checker and calculating the uncertainty using the VGC. In the case of using a VGC, the values of the error factors are given randomly under a specific distribution according to the characteristic of each error factor. Many measurement results are obtained by repeating virtual measurements under this condition. The distribution of these measurement results is analyzed, the variability of this distribution is calculated, and the uncertainty is calculated according to equation (1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003412_j.issn.1004-4132.2011.04.017-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003412_j.issn.1004-4132.2011.04.017-Figure4-1.png", + "caption": "Fig. 4 Topology 1", + "texts": [ + "3, we find that each agent converges to the leader\u2019s states with the initial states of the agents generated randomly and the static leader\u2019s position chosen as x0 = 3 (see Fig. 3). However, we find that the LMI (9) has no feasible solutions for the large enough control gain, i.e., the control gain k can not be too large. Example 2 Switching interconnection topology. Consider a network of four agents and a static leader given by (10). The interconnection topology of the system is switched between the topology 1 in Fig. 4 and the topology 2 in Fig. 5. we choose \u03b31 = 0.03, \u03b32 = 0, \u03b33 = 0.02, \u03b34 = 0.01, and take k1 = k2 = k3 = k4 = 4 for simplicity. Using the LMI toolbox in Matlab for the LMI (13) in Theorem 3, the maximum communication delay is \u03c4max = 2.018 s, i.e., the LMI (13) holds for each topology if \u03c4 2.018 s. In the simulation, the interconnection topology is switched from one topology to another one every 3 s. With the initial states of the agents generated randomly, the static leader\u2019s position is chosen as x0 =3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000951_2009-01-1465-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000951_2009-01-1465-Figure7-1.png", + "caption": "Figure 7: Single Cavity Supercharger Variable Drive", + "texts": [ + " This provides the potential envelope of supercharger operating speeds as shown in Figure 6. With a clear focus on achieving a low cost, weight and package solution, a variable drive for a supercharger needs to be based upon the two roller \u2018yoke\u2019 control Variator design from the OPE transmission. Two supercharger drive designs have been developed in single and twin cavity format with a high level of commonality between designs with both employing 50mm diameter rollers. The single cavity design is shown in Figure 7 : For a traction drive device to operate, the discs and rollers need to be clamped together. Various method of achieving the necessary clamping or \u201cEnd load\u201d force exist from complex hydraulic arrangements to simple bevel spring designs. Focussing on cost, the Variator geometry and associated gearing have been designed in conjunction with the speed and load characteristics of a supercharger to produce a relatively flat high reaction force End Load requirement curve (Figure 8). Therefore, application of a simple bevel spring arrangement to generate the clamping forces together with a thrust bearing to react the loads is possible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003879_j.ymssp.2013.04.009-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003879_j.ymssp.2013.04.009-Figure1-1.png", + "caption": "Fig. 1. Lumped-parameter model of a tuned planetary gear. The fe1 ; e2 ; e3g basis is attached to the carrier and rotates at the constant speed \u03a9c . The stationary basis fE1;E2 ;E3g is not shown.", + "texts": [ + " This includes free response and forced response near resonance where a single mode dominates. A stationary observer measures gear motions by fixed displacement probes and lasers. Accelerometers or strain gauges mounted to the individual gear bodies give the response seen by a rotating observer, therefore, they will have similar frequency content. The results identify the frequency content differences that occur for these two measurements. The tuned (all planets are identical), lumped-parameter model of a spur planetary gear is shown in Fig. 1. The nondimensional matrix equation of motion for the in-plane, forced, undamped vibration about the non-trivial equilibrium of a planetary gear is [21] M \u20acq \u00fe\u03a9cG _q \u00fe \u00f0Kb \u00fe Km\u2212\u03a92 cK\u03a9\u00deq\u00bc f sin \u03b2t; \u00f01a\u00de q\u00bc \u00bdx\u0302c; y\u0302c; u\u0302c; x\u0302r ; y\u0302r ; u\u0302r ; x\u0302s; y\u0302s; u\u0302s; \u03b6\u03021; \u03b7\u03021; u\u03021;\u2026; \u03b6\u0302N ; \u03b7\u0302N ; u\u0302N T ; \u00f01b\u00de where the hats denote non-dimensional physical coordinates. The planetary gear model and the various reference frames (referred to as bases in this work) are described in detail in Ref. [17]. All matrices in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003733_ecc.2013.6669269-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003733_ecc.2013.6669269-Figure2-1.png", + "caption": "Fig. 2. The ball and plate system", + "texts": [ + "20 seconds, an overshoot approximately equal to 5 percent. Each of them are connected to the side of the plate, by using two DOF gimbals. The sampling time of the system and the frame rate provided by the camera are 1 millisecond and 60 frames per second, respectively. Hence, the image information is 978-3-033-03962-9/\u00a92013 EUCA 2855 renewed every about 17 milliseconds. Moreover, the visual system has a constant time delay less than 60 milliseconds. The X-direction of the ball and plate system is illustrated in Fig. 2. Assume that the ball is completely symmetric and homogeneous, and no slipping on the plate. Assume also that all frictions are neglected. The plate rotates in the XYCartesian coordinates with the origin at the center of plate. The equations of motion are as follows: ( mb + Ib r2b ) x\u0308b \u2212mb ( xb\u03b1\u03072 + yb\u03b1\u0307\u03b2\u0307 ) +mbg sin\u03b1 = 0, ( mb + Ib r2b ) y\u0308b \u2212mb ( yb\u03b2\u03072 + xb\u03b1\u0307\u03b2\u0307 ) +mbg sin\u03b2 = 0, where (xb, yb) is the position of the ball on the plate, \u03b1 and \u03b2 are inclination angle of plate to X and Y axis, respectively, mb is the mass of the ball, rb is the radius of the ball, g is gravitational acceleration, Ib is the inertia of the ball, Ltbl is the side length of the plate, rarm is the length between the joint and the center of the load gear" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002591_10402004.2011.626144-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002591_10402004.2011.626144-Figure10-1.png", + "caption": "Fig. 10\u2014Effect of combined influence of surface roughness and nonNewtonian behavior of lubricant on journal center trajectory.", + "texts": [ + ", when M\u0304j = 1.5792), the size of the journal center whirl orbit is observed to be almost same for both longitudinal and transverse roughness patterns. When the journal mass is further increased four times (i.e., when M\u0304j = 3.1583), the transverse roughness pattern is observed to make the journal center whirl orbit smaller than that of the longitudinal pattern. From Fig. 9 it can also be observed that as the value of the journal mass increases, the size of the journal center whirl orbit increases. Figure 10 shows the combined influence of surface roughness and non-Newtonian behavior of a lubricant on journal center motion trajectory. From Fig. 10 it can be observed that the difference between the size of journal center whirl orbits of smooth and rough bearings (moving roughness with transverse D ow nl oa de d by [ N ew Y or k U ni ve rs ity ] at 1 6: 43 0 5 O ct ob er 2 01 4 roughness pattern) is large when surface roughness and nonNewtonian behavior of the lubricant effects are considered together, whereas it is comparatively smaller when the surface roughness effect alone is considered. These results clearly indicate an interactive influence of the surface roughness and nonNewtonian behavior of the lubricant on the bearing response" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure10-1.png", + "caption": "Fig. 10 Metamorphic mechanism with Link 4 blocked by pin P2", + "texts": [], + "surrounding_texts": [ + "f t\nm m i p D l a s n\nw e m b t t\nm a c d c t r\nF m\nF c\n0\nDownloaded Fr\nA2 = 0 \u2212 1 0 0 1 \u2212 1 0 1 0 0 0 1 0 1 0\n0 0 1 0 1 1 0 0 1 0 4\nNote that A1 and A2 are different. Moreover, the complete inormation of the topological change is kept in the adjacency marix.\nThe proposed method can be used to compute the DOF of the etamorphic mechanisms as well. The adjacency matrix of a etamorphic mechanism is the same as the ordinary mechanisms f there is no \u22121 element in the matrix. In this case, the metamorhic mechanism behaves just like an ordinary mechanism with OF Fm 2, where Fm is the maximal DOF. When some of the inks are attached together, the number of effective links reduces nd the DOF of the mechanism decreases as well. Let us consider imple spherical linkage mechanisms and planar linkage mechaisms. In this case, the DOF of a mechanism can be calculated as\nF = 3 n \u2212 1 \u2212 2p 5\nhere n denotes the number of effective links and p the number of ffective kinematic pairs. Suppose the number of links of a metaorphic mechanism is n0, the size of the adjacency matrix should e n0 n0. Examining the upper or lower triangle elements of he adjacency matrix and denoting the number of \u22121 as n\u22121 and he number of 1 as n1, we have the following relations:\nn = n0 \u2212 n\u22121 6\np = n1 7 Moreover, all the possible configurations of a metamorphic echanism can be found. Suppose Fm=2, as the example given\nbove, then changing any pair of elements of the original adjaency matrix symmetrical from value 1 into value \u22121 will give a ifferent configuration of the metamorphic mechanism. If Fm=3, hanging any two pairs of elements of the original adjacency marix from value 1 into value \u22121 will also give a different configuation of the metamorphic mechanism, and so on.\nig. 1 Illustration of five-bar spherical metamorphic echanism\nig. 2 The configuration when a five-bar linkage configuration hanges to a four-bar linkage configuration\n74501-2 / Vol. 130, JULY 2008\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash\n3 Examples of Planar Mechanisms Figure 3 shows a planar mechanism with five links. The original adjacency matrix of the mechanism, A0, is the one in Eq. 1 with Fm=2. There are five more possible configurations of the mechanism as there are five 1 among upper triangle of A0. Each one will result in a different adjacency matrix as shown below.\n1 If Slider 2 is fixed to Link 1, as shown in Fig. 4 a , the prismatic pair between Link 1 and Slider 2 is frozen. The mechanism is termed as Configuration 1 and is shown in Fig. 4 b . It works as a general four-bar linkage mechanism with DOF F=1. The adjacency matrix A1 can be obtained by multiplying \u22121 on A0 1,2 and A0 2,1 , i.e.,\nA1 = 0 \u2212 1 0 0 1 \u2212 1 0 1 0 0 0 1 0 1 0\n0 0 1 0 1 1 0 0 1 0 8\n2 If Slider 2 is attached together with Link 3 by a pin, P, as shown in Fig. 5 a , the revolute pair between Slider 2 and Link 3 is then frozen. The mechanism is termed as Configuration 2 and is shown in Fig. 5 b . Its adjacency matrix, A2, can be obtained by multiplying \u22121 on A0 2,3 and A0 2,3 , i.e.,\nFig. 5 Configuration 2 of the example\nTransactions of the ASME\nx?url=/data/journals/jmdedb/27877/ on 02/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "J Downloaded Fr\nA2 = 0 1 0 0 1 1 0 \u2212 1 0 0 0 \u2212 1 0 1 0\n0 0 1 0 1 1 0 0 1 0 9\nIt is interesting to know that if this configuration comes from Configuration 1, the adjacency matrix A2 can be obtained by multiplying \u22121 on A1 1,2 , A1 2,1 , A1 2,3 , and A1 3,2 . Moreover, according to Eqs. 5 \u2013 7 , n=n0 \u2212n\u22121=5\u22121=4, p=n1=4, and F=3 n\u22121 \u22122p=3 4\u22121 \u22122 4 =1. 3 If Link 3 is attached together with Link 4 by a pin, P, as shown in Fig. 6 a , the revolute pair between Link 3 and Link 4 is then frozen. The mechanism is termed as Configuration 3 and is shown in Fig. 6 b . If this configuration comes from Configuration 2, the adjacency matrix A3 can be obtained by multiplying \u22121 on A2 2,3 , A2 3,2 , A2 3,4 , and A2 4,3 , i.e.,\nA3 = 0 1 0 0 1 1 0 1 0 0 0 1 0 \u2212 1 0\n0 0 \u2212 1 0 1 1 0 0 1 0 10\n4 If Link 4 is attached to Link 5 by a pin, P, as shown in Fig. 7 a , the revolute pair between Links 4 and 5 is then frozen. The mechanism is termed as Configuration 4 and is shown in Fig. 7 b . If this configuration comes from Configuration 3, the adjacency matrix A4 can be obtained by multiplying \u22121 on A3 3,4 , A3 4,3 , A3 4,5 , and A3 5,4 , i.e.,\nA4 = 0 1 0 0 1 1 0 1 0 0 0 1 0 1 0\n0 0 1 0 \u2212 1 1 0 0 \u2212 1 0 11\nournal of Mechanical Design\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash\n5 If Link 1 is attached together with Link 5 by a pin, P, as shown in Fig. 8 a , the revolute pair between Links 1 and 5 is frozen. The mechanism is termed as Configuration 5 and is shown in Fig. 8 b . If this configuration comes from Configuration 4, the adjacency matrix A5 can be obtained by multiplying \u22121 on A4 4,5 , A4 5,4 , A4 1,5 , and A4 5,1 , i.e.,\nA5 = 0 1 0 0 \u2212 1 1 0 1 0 0 0 1 0 1 0\n0 0 1 0 1 \u2212 1 0 0 1 0 12\nNote that among all possible configurations, any one can be transformed to any other one. Though, the adjacency matrix can always catch the process of the change. Figure 9 shows such an example. This is a planar metamorphic mechanism with five links. There is a spring embedded in Link 1, which can push Slider 2 moving along the slot. There are two pins, P1 and P2, on the frame to limit the swing of Link 4. When Link 4 is blocked by one of the pins P1 or P2, Slider 2 will slide along the slot of Link 1, as shown in Figs. 10 and 11. The mechanism works as a guide-bar mechanism in these cases.\nHere, only two configurations are displayed: If Link 4 does not touch both pins P1 and P2, Slider 2 will rest at the end of the slot\nJULY 2008, Vol. 130 / 074501-3\nx?url=/data/journals/jmdedb/27877/ on 02/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "f l L F a A 1 p p c\np f a m\nF\n0\nDownloaded Fr\norced by the spring. At this time, the mechanism is a four-bar inkage mechanism as that of Configuration 1 with F=1. When ink 4 is blocked by either P1 or P2, Configuration 4 appears with =1 and the adjacency matrix is A4. The transformation from A1 nd A4 can be done by multiplying \u22121 on A1 4,5 , A1 5,4 , 1 1,2 , and A1 2,1 , and vice versa. One can imagine when Link continuously rotates, two different configurations alternately apear and Link 4 swings with dwells at both limiting positions. The resented new adjacency matrix can describe the topological hanges of this mechanism effectively.\nFrom a practical point of view, the aforementioned metamorhic mechanism can replace a cam mechanism with oscillating ollower in some sense and the swinging range of Link 4 can be djusted by the positions of the two pins. The details of the echanism will be the topic of another study.\nig. 11 Metamorphic mechanism with Link 4 blocked by pin P1\n74501-4 / Vol. 130, JULY 2008\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash\n4 Conclusions This short paper introduces a new adjacency matrix for the metamorphic mechanism. Based on the discussion above, the following conclusions can be drawn.\n1 By introducing a \u22121 for the fixed kinematic pair of a metamorphic mechanism, the new adjacency matrix can effectively describe configurations of metamorphic mechanisms. The new adjacency matrix always has the same dimension and can effectively trace the structural changes of the metamorphic mechanism. 2 The new adjacency matrix can be used to find all the possible configurations and to calculate the DOF of the metamorphic mechanism.\nAcknowledgment This paper is partially supported by a grant from Hong Kong Innovation and Technology Fund No. ITS/001/05. This work was completed when Z. H. Lan worked in the Institute of Precision Engineering, Chinese University of Hong Kong.\nReferences 1 Dai, J. S., and Rees, J. J., 1999, \u201cMobility in Metamorphic Mechanism of\nFoldable/Erectable Kinds,\u201d ASME J. Mech. Des., 121, pp. 375\u2013382. 2 Tsai, L. W., 2001, Mechanism Design: Enumeration of Kinematic Structures\nAccording to Function, CRC,Boca Raton, FL. 3 Dai, J. S., and Rees, J. J., 2005, \u201cMatrix Representation of Topological\nChanges in Metamorphic Mechanisms,\u201d ASME J. Mech. Des., 127, pp. 837\u2013 840. 4 Dai, J. S., Ding, X., and Wang, D., 2005, \u201cTopological Changes and the Corresponding Matrix Operations of a Spatial Metamorphic Mechanism,\u201d Chin. J. Mech. Eng., 41 8 , pp. 30\u201335. 5 Dai, J. S., and Zhang, Q. X., 2000, \u201cMetamorphic Mechanisms and Their Configuration Models,\u201d Chin. J. Mech. Eng., 13 3 , pp. 212\u2013218.\nTransactions of the ASME\nx?url=/data/journals/jmdedb/27877/ on 02/05/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use" + ] + }, + { + "image_filename": "designv11_3_0003449_imece2011-63452-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003449_imece2011-63452-Figure3-1.png", + "caption": "FIGURE 3: Frequency-response-function with (A,B,C,D) four dominant mode-shapes for \u03b8 = \u03c0 2 , \u03b8 \u2217 = \u03c0 4 and \u03b3 =\u2212\u03c0 4 .", + "texts": [ + " Substitution of q = U\u03b7 and F = UQ into eq. 21 and premultiplication with UT gives an uncoupled equation in terms of generalized coordinates \u03b7 . \u03b7\u0308(t)+ [\u2126]\u03b7 = Q (23) where [\u2126] = diag(\u21262 1, \u00b7 \u00b7 \u00b7 ,\u21262 N) and Q is the generalized force matrix. From 23, generalized equation for rth coordinate, with modal damping factor \u03ber, can be written as \u03b7\u0308r +2\u03ber\u2126r\u03b7\u0307r +\u2126 2 r \u03b7r = Qr (24) Now consider a defect located at an angle \u03b3 in rotating CS on inner or outer race of the planetary bearing with position angle \u03b8 \u2217 (fig. 3). Due to the dynamic interaction between defect and rolling-elements, a force F\u03b4 is applied to both ith planet gear (i = \u0393) and bearing inner-race in opposite directions. Therefore, Fpi = F\u03b4 (cos\u03b3,sin\u03b3,0)T (25a) Fb =\u2212Fpi . (25b) Using Q = UT F, generalized force for rth coordinate can be calculated as Qr = ( u(r) pi )T Fpi + ( u(r) b )T Fb = (u(r)pbx cos\u03b3 +u(r)pby sin\u03b3)F\u03b4 (26) where u(r) pi = (u(r)pix ,u(r)piy ,u(r)pi\u03b8 )T , u(r) b = (u(r)bx ,u(r)by ,u(r)b\u03b8 )T and u(r)pb(\u2022) = u(r)pi(\u2022) \u2212u(r)b(\u2022) ", + " 29 and 2, ring gear response in the radial and tangential directions are obtained and corresponding frequency-response-functions (FRFs) are also calculated. FRF for ring gear radial response is given below. H(\u03b3,\u03b8 \u2217,\u03b8 ,\u03c9) = W\u0304r(\u03b8 ,\u03c9) F\u03b4 (\u03c9) = \u2211 \u03a6 n=0{a\u0304n(\u03c9)Cn\u03b8 + b\u0304n(\u03c9)Sn\u03b8} F\u03b4 (\u03c9) = N \u2211 r=1 u(r)pbx cos\u03b3 +u(r)pby sin\u03b3 \u21262 r \u2212\u03c92 +2\u03b9\u03ber\u03c9\u2126r \u03a6 \u2211 n=0 [{Cn\u03b8Cn\u03b8\u2217 +Sn\u03b8 Sn\u03b8\u2217}u(r)an +{Sn\u03b8Cn\u03b8\u2217 \u2212Cn\u03b8 Sn\u03b8\u2217}u(r)bn ] (30) FRF given by eq. 30 is the function of fault position angle (\u03b3), carrier rotation angle (\u03b8 \u2217), and position angle of measurement point (\u03b8 ). Figure 3 shows the frequency-response-function for \u03b8 = \u03c0 2 , \u03b8 \u2217 = \u03c0 4 , \u03b3 =\u2212\u03c0 4 and \u03ber = 0.05 \u2200 r. Model parameters are listed in table 1. In all the dominant modes (A, B, C and D), ring gear doesn\u2019t behave as a rigid body. Therefore, flexibility of the ring is important for response prediction. According to eq. 30, if both ur pbx = 0 and ur pby = 0, i.e. planet gear and bearing inner-race are moving as a rigid body, FRF will be zero. Figure 3 also shows that the dominant modes have relative displacement between planet and bearing inner-race. 6 Copyright c\u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use The impact force produced when rolling-elements strike the fault is modelled as an impulse train shown in figure 4. The frequency of the impulse train \u03c9d is the characteristic fault frequency and its value depends on the bearing geometry and operating speeds (see Appendix)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002736_tmag.2012.2196502-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002736_tmag.2012.2196502-Figure8-1.png", + "caption": "Fig. 8. Distributions of the maximum flux density , the maximum field strength and the iron loss in the PM motor model core.", + "texts": [ + " The rolling direction agrees with the X-direction. However the local easy axis depends on the residual stress distribution. The compressive stress in R.D. and T.D. was distributed mainly in the back yoke and the largest tensile and compressive stress occurred around the caulking. Therefore, it can be considered that the local magnetic properties in these areas are deteriorated. Because the magnetic path is confined due to the residual stress, the magnetic flux will concentrate in a local area. Fig. 8 shows the analyzed results of , and distributions with and without stress. The values around the slots and the back yoke part were large under the stressed condition, and those were larger than that without stress. It can be also seen that values around caulking area became small due to deterioration of the magnetic property caused by the compressive stress around caulking. under stress in the back yoke was much larger than that without stress. The cause can be considered that the residual stress in the right hand side back yoke is almost compressive stress in both R" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001707_bsn.2010.46-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001707_bsn.2010.46-Figure3-1.png", + "caption": "Fig. 3. Side view of club face and ball at impact showing loft angle.", + "texts": [ + " While examining the figure note the coordinate system that will be used throughout this paper. Positive z points upward perpendicular to the ground, positive x denotes the intended ball path and positive y points to the golfer while remaining perpendicular to x and z. There are many parameters of a golf swing that affect the trajectory of the golf ball. The goal of our model is to identify parameters that address precision and repeatability of the swing. We focus our attention on the following most critical parameters: \u2022 Face angle (\u03c8) at impact (Figure 2) \u2022 Loft angle (\u03b8) at impact (Figure 3) \u2022 Lie angle (\u03c6) at impact (Figure 4) \u2022 Velocity throughout swing \u2022 Location (x\u2032, y\u2032) of impact on club face (Figure 4) \u2022 Motion path immediately surrounding impact1 \u2022 Tempo: Proportion of back-swing duration to forwardswing duration By examining each of Figures 2, 3 and 4 it can be observed that the most predictable path of travel for the golf ball will occur when loft, face and lie angles are all zero. In fact the motion is so sensitive to error that a \u03c8 = \u00b13 degree face angle will result in an error greater than 15cm for a 3m putt" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003133_j.jmrt.2013.10.002-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003133_j.jmrt.2013.10.002-Figure4-1.png", + "caption": "Fig. 4 \u2013 Electron beam melting (EBM) system schematic.", + "texts": [ + " Special nterest involved the observations of the precursor powder icrostructure and the EBM-fabricated components along ith their residual tensile and hardness properties. Reticuated mesh components were also fabricated by EBM, and he Young\u2019s modulus (dynamic stiffness) measured for these omponents with varying density (or porosity) using resonant requency detection. . Experimental procedures apidly solidified, gas atomized pure iron powder having izes, size distribution, and microstructures as shown in igs. 1\u20133 were processed in an upgraded, Arcam S-12 electron eam melting (EBM) system illustrated schematically in Fig. 4. igs. 1(a) and 3(a) show the details of the precursor powder orphologies and size distribution, with an average particle ize of 19 m as shown in Fig. 3(a). Fig. 1(b) shows polished nd etched particle cross-sections illustrating an irregular but quiaxed grain structure of roughly 3 m diameter as illusrated in the magnified scanning electron microscope (SEM) mage in Fig. 2. Although the powder crystal structure exhibted primarily -Fe (bcc: a = 2.87 A\u030a), a -Fe (1 1 0) (bcc: a = 3.0 A\u030a) eak was observed by X-ray diffraction (XRD) analysis as indiated by the arrow in Fig. 3(b). The EBM system (Fig. 4) upgrade consisted of an air-cooled lectron gun (at G in Fig. 4), with an insulating space between at scan speeds of 200\u2013300 mm/s at increased beam current. The system operates at a beam voltage of 60 kV in a vacuum of \u223c104 Torr. In this research program, solid cylindrical components were fabricated measuring 1.5 cm diameter \u00d7 10 cm in length, along with a range of reticulated mesh structures using a dode-thin build element (a rhombic dodecahedron shape) using Materialise Software, embedded in a CAD program to direct the layer melting. This mesh element was selectively manipulated and expanded to produce open-cellular mesh components having different strut diameters (ranging from \u223c1 to 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000826_j.euromechsol.2009.07.005-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000826_j.euromechsol.2009.07.005-Figure8-1.png", + "caption": "Fig. 8. {X(u)}{X 3(v)} motion generators with two Pas and one P pair.", + "texts": [], + "surrounding_texts": [ + "In what follows, we will briefly recollect X motion serial generators enumerated in (Herve\u0301 and Sparacino, 1991; Lee and Herve\u0301, 2005b, 2009). As we know, the serial setting of two kinematic pairs produces a kinematic bond between the distal bodies of the obtained chains including three bodies and this bond is the product of the pair bonds. This is why the closure of the product in a subgroup of a group plays a key role. As a consequence, a Schoenflies group can be generated by many serial arrays of kinematic pairs. The 4D group {X(u)} can be decomposed into a product of four 1D subgroups associated with the 1-dof Reuleaux pairs (Reuleaux, 1875). A generic example of such a product is {H(N1, u, p1)}{H(N2, u, p2)}{H(N3, u, p3)}{H(N4, u, p4)}. It is possible to replace one, two or three of the factors {H(Ni, u, pi)} by a 1D subgroup {T(si)} of translation parallel to the unit vector si, provided that the vectors si are linearly independent. One, two or three pitches may also be equal to zero. Obviously, the above expression is valid in a neighborhood of the identity E if and only if (iff) the product is a 4D manifold, or else the product as well as its corresponding chain is singular. When the above expression is not valid for any value of parameters, singularity may occur on the frontier of the neighborhood. Any serial arrangement of four 1-dof kinematic pairs without intermediate link having redundant passive motion makes up a mechanical generator of the subgroup {X(u)}. The comprehensive list of the general-type concatenations of 1-dof kinematic pairs generating Schoenflies motion is shown in Table 1. The total number of these generators is fifteen. However, when R pairs replace one or several H pairs, there are forty-three general architectures of X motion generators. Moreover, these combinations can be sorted into four classes based on the number of prismatic pairs (Lee and Herve\u0301, 2005b, 2009). The most typical representative of the primitive generators of an X-group is HHHH where the screws H have parallel axes and the four pitches must not be all equal. A P pair is a limit case of an H pair either with a pitch becoming infinite or with an axis going at infinity, or a combination of both previous situations. An R pair is an H with a zero pitch. All the corresponding kinematic chains that generate X-motion are graphically demonstrated in (Lee and Herve\u0301, 2005b, 2009). Clearly, mechanical generators of X motion with a parallel or hybrid topology also exist and some are described in the literature (Angeles, 2004; Gogu, 2007; Salgado et al., 2007). The planar hinged parallelogram produces circular translation between two opposite bars. The corresponding composite joint is denoted as Pa. Circular translation is a 1D submanifold of an X group. Consequently, Pa, as shown in Table 2 where some generators having the equivalent architectures by kinematic inversion are cancelled out can replace the P pairs in the combinations of Table 1." + ] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure1-1.png", + "caption": "Fig. 1. Spherical bushing joint.", + "texts": [ + " Consequently, the fulfillment of the kinematic constraints is more easily achieved and the size of the integration time steps, selected by the numerical integrator of the equations of motion, used during the numerical integration is larger leading to faster computation times. However, for stiffer bushings, or for compliances with particular nonlinear constitutive laws, this trend may not apply due to the high frequency contents that they bring into the multibody system response. The spherical joint allows for three free rotations between the bodies connected, penalizing only the relative translation displacements. A set of translational springs is used to represent the spherical bushing and limit the relative motion of the bodies connected, as presented in Fig. 1. The bushing reaction forces over bodies i and j are represented as 1[ ( ) ( ) ]i j i K f b d \u03b4 \u03b4 \u03b4 \u03b4= \u2206 + + = \u2212 f d f f & (3) where ( )K \u03b4 \u03b4\u2206 is the spring nonlinear elastic force increment, f(\u03b4) represents the equivalent spring force due to its state of deformation, b is the stiffness proportional damping, and \u03b4 is the bushing deformation. Vector d, is the distance between Pi and Pj, as P P j j i i= + \u2212 \u2212d r s r s (4) and d is the length of vector d given by Td = d d while\u03b4& is the time rate of change of\u03b4 . Note that the difference between d and \u03b4 is simply a shift that accounts for the undeformed thickness of the bushing. When no gap exists between the two bodies involved in the joint d=\u03b4. Assuming that no gap exists in the joint, d=\u03b4, and, according to Fig. 1 the time derivative of \u03b4 is T \u03b4 \u03b4 = d d&& (5) where d& is the time derivative of Eq. (4) as P P j j j j i i i i= + \u2212 \u2212d r A \u03c9 s r A \u03c9 s& & % & % (6) where \u03c9 is the angular velocity of body i. When the bushing joint is not located in the center of mass of the connected bodies the transport moments are ( ) ( ) ' ' T P i i i i T P j j j j = = n A s f n A s f % % (7) where P is% is a skew-symmetric matrix made with the components of vector P is . The degrees of freedom to be penalized by this joint are the normal translational displacement, dn, depicted in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000871_09544062jmes949-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000871_09544062jmes949-Figure3-1.png", + "caption": "Fig. 3 FEM of the HD", + "texts": [ + "comDownloaded from HD behaviour, given that the aforementioned (twodimensional) methodology did not take into account the complex deformation of the FS under loading conditions. Each part of the HD undergoes some degree of modelling in order to replicate the torsional stiffness of the actual HD and to minimize computational time. The ANSYS software is used as it has proved to be effective in simulating complex contact problems [13]. Geometrical dimensions were measured in the laboratory for this study. All components presented in this article are equipped with a structured mesh to improve accuracy and to reduce computational time [14]. Figure 3 shows the model of the HD created using the gathered data. It contains \u224860 000 uniaxial, brick, shell, and contact elements. The following sectionspresent thecomponentsof theHDand their FEMs. Even though the WG shown in Fig. 1 seems to be a fairly rigid structure, it ismodelled taking into account the radial forces that are acting on the ball bearings. The WG is modelled through three distinctive parts: an inner and an outer ring and uniaxial elements; the inner and outer rings are made of brick elements connected through compression-only uniaxial elements (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000826_j.euromechsol.2009.07.005-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000826_j.euromechsol.2009.07.005-Figure11-1.png", + "caption": "Fig. 11. In-parallel layout of chains generating the same X\u2013X motion.", + "texts": [ + " When discriminating R pairs that are H pairs with a zero pitch and H pairs with a non-zero pitch, the previous eight general chains become 106 distinct generators. In addition, nineteen more realizations of {X(u)}{H(M, v, q)} are X\u2013X motion generators with at least one hinged parallelogram, which leads to 148 combinations when an arbitrary number of R pairs substitutes for the corresponding number of H pairs. Finally, one can notice that if a fixed base is connected to a moving platform by two limbs, each limb generating a given specified X\u2013X motion, then the platform motion is the same X\u2013X motion. A typical example is depicted in Fig. 11. This closed-loop chain is movable with 5 degrees of freedom while the Chebyshev\u2013 Gru\u0308bler\u2013Kutzbach formula for spatial linkages yields 4 degrees of freedom. Such a result is not trivial. Obviously, the X\u2013X motion generators that are disclosed in the paper can be employed in the synthesis of sub-6-dof parallel manipulators. The authors are very thankful to the National Science Council for supporting this research under grant NSC 95-2221-E-151-011 and NSC 98-2221-E-151-019. Angeles, J., 2004. The qualitative synthesis of parallel manipulators" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003078_j.jmps.2011.10.003-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003078_j.jmps.2011.10.003-Figure3-1.png", + "caption": "Fig. 3. Illustration of the s-jump rule Eq. (17).", + "texts": [ + " Indeed, Hs c\u00f0r\u00de depends only on the stick\u2013slip boundary s and contact radius c, and does not depend on how this situation has been obtained. The reversible solution, on the contrary, contains the path-integral by way of c(T) or N(T). An important property of the hysteretic solution which is closely related to the path-invariance is that for srxrc Hs c\u00f0r\u00de \u00bcHs x\u00f0r\u00de\u00feHx c \u00f0r\u00de: \u00f017\u00de This identity can be easily derived from Eq. (4), and can be interpreted in the following manner. Suppose that the application of a given normal and tangential force on an initially traction-free contact has created a distribution Hs x\u00f0r\u00de (see Fig. 3). At first, compression is increased at constant tangential force, so that the contact radius grows from x to c. Next, the tangential force is increased at constant compression so that the stick\u2013slip boundary starts to shrink departing from c and finally reaching the value of x. The final traction distribution Hs x\u00f0r\u00de\u00feHx c \u00f0r\u00de then fully coincides with Hs c\u00f0r\u00de. The latter distribution satisfies the condition Hs c\u00f0r\u00de \u00bc ms\u00f0r\u00de in the range srrrc, which means that the actual stick\u2013slip boundary for the traction distribution Hs x\u00f0r\u00de\u00feHx c \u00f0r\u00de equals s", + " The nonhysteretic deformation of the spheres with s\u00bcc ends at point PR\u00feH when slip again starts propagating from the contact radius inward. At a specific point, labeled D in Fig. 5, the stick\u2013slip boundary s reaches the last memory point s3. Fig. 6(d) visualizes what happens at that instance. The reversible solution for traction becomes fully compensated, which means that curve CO in Fig. 6(d) equilibrates the component WR in Fig. 6(b). The stick\u2013slip boundary s instantly jumps from s3 to s2, similar to the situation discussed in Fig. 3, and at even later stages s continues to propagate inward. This means that the two memory points s2 and s3 are erased, and the system further evolves according to the solution Eqs. (19)\u2013(21) as if the reversible regime has never occurred. Indeed, by substituting s\u00bcs3 c2 and t3(r) t2(r) into Eq. (24), we obtain that t\u00f0r\u00de \u00bc t2\u00f0r\u00de\u00feRc2 c2 \u00f0r\u00de\u00feHc2 c \u00f0r\u00de, where Rc2 c2 \u00f0r\u00de is zero. The contribution t2(r) should be taken from Eq. (19) with s\u00bcs2 and c\u00bcc2, with the result t\u00f0r\u00de \u00bc t1\u00f0r\u00de\u00feHs2 c1 \u00f0r\u00de\u00feHs2 c2 \u00f0r\u00de\u00feHc2 c \u00f0r\u00de: Taking into account the s-jump rule, Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001515_09596518jsce826-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001515_09596518jsce826-Figure1-1.png", + "caption": "Fig. 1 XY fine positioning stage", + "texts": [ + " To measure and manipulate structures on the nanometre scale, highresolution positioning stages are used. These stages are able to position a pattern in all three dimensions with an accuracy of less then one nanometre with operational ranges up to several hundred millimetres. The work presented here is motivated by a 2006200mm2 fine positioning stage, which was developed within the Collaborative Research Centre 622 \u2018Nanopositioning and Nanomeasuring Machines\u2019 at the Ilmenau University of Technology [1, 2]. As can be seen in Fig. 1, each axis is driven by two linear voice coil actuators. The actuators are powered by proprietary analogue amplifiers, which provide the necessary current with the required precision. Each axis is supported by two linear Vgrooved guideways. The position is measured by stabilized HeNe laser interferometers with a resolution of less than 0.1 nm [3]. For data acquisition and control, a modular dSpaceH real-time system in combination with MATLAB/SimulinkH is utilized [4]. The control algorithm works with a sample rate of 10 kHz and operates on the amplifiers with a 16-bit resolution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001547_iciea.2009.5138555-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001547_iciea.2009.5138555-Figure2-1.png", + "caption": "Fig. 2. Approximate waveforms of stator currents and back-emfs for conventional commutation.", + "texts": [ + " The inverter power switches are generally controlled using pulse width modulation (PWM) patterns. If a fast inner control loop is applied to the drive such as the current control loop or the direct torque control loop, the PWM duty cycle is controlled in one switching cycle to achieve demanded control goal. In addition, the inverter gating signal patterns can be properly selected such as the soft chopping patterns [8]. However, because of the inductive stator winding time constant, the actual stator current lags the demanded current as shown in Fig. 2. Hence, a phase advanced commutation scheme is desired to modify the actual current waveform to improve the drive performance as shown in Fig.3. As compared to the BLDC motor, the IPM-BLDC motor drive needs a larger phase advanced angle to result in the stator current with a larger leading angle such that a better performance could be expected. The detailed features will be discussed in the next section. As reference to Fig. 1 is the power stage of a three phase BLDC motor drive. In the BLDC motor model, SR and SL are the equivalent phase winding resistance and decoupled phase equivalent inductance respectively. Three phase back-emfs are denoted as ane , bne and cne respectively. The inverter power switches are controlled by the soft chopping patterns [8]. The proper switching patterns are generated according to the combination of signals from three rotor position hall sensors. There are six combinations in one electric period and hence there are six different switching patterns. Considering the typical waveforms of conventional commutation as shown in Fig. 2, taking the commutation time period between \u201cS3 and S6 PWM\u201d and \u201cS2 and S3 PWM\u201d as an explanation example, phase current Ai starts increasing toward the desired current amplitude as well as phase current Ci starts decreasing toward zero. The corresponding phase A and C voltage equations can be expressed as anBnAB A SAS evV dt di LiR -+=+ (1) cnBnCB C SCS evV dt di LiR -+=+ (2) where ABV and CBV are two set of line to line voltages , Bnv is motor phase B voltage. At the commutation time period, as S3 and S2 are turned on, ABV equals the DC link voltage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001663_s12239-009-0049-6-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001663_s12239-009-0049-6-Figure4-1.png", + "caption": "Figure 4. Vehicle coordinate system with regard to the kingpin axis.", + "texts": [ + " When this steering rack force opposes a driver\u2019s steering input and is capable of overcoming steering friction (steering gear box friction, strut bearing friction, and tire friction), the steering wheel automatically restores to the straight-ahead position. In this section, we propose an analytic model that considers returnability with an analysis of steering rack force for each directional tire force (vertical, lateral, and longitudinal). 4.1. Vertical Force A local coordinate system that follows the vehicle coordinate system is defined on the tire contact patch plane with origin A (contact patch point) in Figure 4, where D is the point where the kingpin axis crosses the contact patch plane, rc is the kingpin offset at the ground, nc is the caster trail, and \u03b4 is the tire turning angle. Moment (MV) by vertical force is defined as the cross-product of the displacement vector (rA) and the vertical axle force (FV). (1) , where As the weight of a vehicle body transfers due to lateral acceleration, the vertical axle load changes. This means that the inner axle load decreases and the outer axle load increases, even though lateral acceleration is less than 1", + " Camber angle with respect to the road varies over the range 0o ~ \u22121o at the outer road wheel and +5o ~ +7o at the inner wheel. When there is negative camber at the outer front wheel and positive camber at the inner front wheel, the lateral forces of both wheels are in the direction of the turning center. (inner wheel) (8) (outer wheel) (9) Equations (10) and (11) describe the moments around the tire contact point for geometric slip and camber angle, respectively. The tire side slip force (FS) acts on point B in Figure 4 with the lever arm (nc + nr), while the camber force (FC) acts on point A with the caster trail lever arm (nc). The kingpin axis moments in Equations (12) and (13) are obtained from the product of the moments (MS and M c ) and the kingpin axis vector (ek). The moment caused by tire side slip force is of consequence in steering returnability, because the caster trail (nc) varies considerably according to the tire turning angle. It means that the outer wheel acts on a smaller caster trail during turning, while the inner wheel acts on a larger caster trail (Figure 3)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002257_j.jsv.2011.06.018-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002257_j.jsv.2011.06.018-Figure1-1.png", + "caption": "Fig. 1. Configuration of the ball joint system.", + "texts": [ + " The contact pressure distribution and the corresponding equilibrium position are numerically found in the balance equation. From the assumed modes method, the equations of motion for the ball joint squeak at the equilibrium are discretized by the system modes. The squeak propensity of the system is then investigated through the eigenvalue sensitivity analysis associated with the design parameters. The ball joint system is modeled as a rotating flexible beam attached to a ball which is in contact with a hemi-spherical socket as shown in Fig. 1. The vibration modes of the socket will not be considered in the current study for the simplification of the problem. Therefore, the socket is assumed to be rigid. This assumption is better for the application where the first natural frequency of the socket is above the audible range [11]. The flexible beam has the constant rotation speed (O), density (r), length (L) and bending stiffness EIzz(x) in the y-direction and EIyy(x) in the z-direction. It is subject to the concentrated inertia of the ball at one end and the free boundary condition at the other end" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003225_s40194-013-0040-8-Figure16-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003225_s40194-013-0040-8-Figure16-1.png", + "caption": "Fig. 16 (continued)", + "texts": [ + " Therefore, even though L is changed, the strength of the interaction near the electrode does not change. However, for a tandem-electrode TIG with a tilted torch (such as in this study), as the arc plasma gets farther from the electrode tip, the arc plasmas get closer to each other. Consequently, the temperature distribution of the arc plasma near the anode metal changes significantly depending on L. In the case of a large L, each arc plasma is easy to hit and deforms more widely. Figure 15 shows the arc shapes in the experiments; they are in very good agreement with the calculation results. Figure 16 shows the numerical results for the weld spots. Figure 16a shows the temperature and velocity distributions on the base metal surface. Comparing i and ii of Fig. 16a, the direction of the major axis is changed. This is because when E is larger, each arc plasma acts as an individual heat source, and so there is less deformation. Focusing on i and iii of Fig. 16a, when L is large, the ellipsoidal shape is more elongated. It can be explained that as L increases, the deformation of the arc plasma becomes wider. Moreover, when both E and L become large, the lengths of the major axis and the minor axis are similar. This result shows that the effects of E and L cancel each other. Focusing on the cross-section of the weld pool, for L=3 and E=10, as shown in ii of Fig. 16b and c, the weld pool is the deepest at the center. In this case, the velocity direction is from the edge to the center in all regions of the weld pool surface. Therefore, when tandem-electrode TIG used, the electrode alignment affects the velocity distribution of the weld pool and changes the cross-sectional shape of the weld pool. The experimental results are show in Fig. 17. These are in very good agreement with the numerical results. In this study, a numerical analysis of a tandem-electrode TIG is conducted using a 3D model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001474_bf00533283-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001474_bf00533283-Figure2-1.png", + "caption": "Fig. 2. Angular momenta of an. element of the shaft", + "texts": [ + " The assumption tha t we have only terminal bearings could be relaxed, but the relaxation would complicate the analysis. The extension of the results we will obtain for this rotor with terminal bearings, to a rotor with any number of arbitrari ly placed bearings is immediate. We shall assume tha t the imposed angular velocity Q is constant. 3. Equations of Motion 3.2. Inert ia forces In order to determine the inertia forces on an element of the shaft we will now consider such an element of length dz in the fixed coordinate system X Y Z (Fig. 2). The deflection of the centre of the element is given by x and y and the angles of rotat ion of the element are given by v and w for rotations in the X Z - and the YZ-planes, respectively. Calculating the moments due to inertia forces we will assume tha t in spite of anisotropy due to unbalance, the equatorial mass moment of inertia can be approximated to half the value of the polar mass moment of inertia of the cross section, which is J = 2 m r 2, where m = re(z) is the mass per Ing.Arch. Bd. 42, H" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.8-1.png", + "caption": "FIGURE 6.8 Typical viscometer.", + "texts": [ + "23 gives the torque per unit length as a2b2 Tq \u00bc 4\u03c0\u03bc \u00f0\u03a9b\u2212\u03a9a\u00de \u00f06:25\u00de b2\u2212a2 Similarly, the torque per unit length for the outer cylinder is given by a2b2 Tq \u00bc \u22124\u03c0\u03bc \u00f0\u03a9b\u2212\u03a9a\u00de \u00f06:26\u00de 2\u2212a2b It should be noted that these equations for the torque per unit length are valid only if the flow remains entirely circumferential. It is possible to make use of Equation 6.26 in the measurement of viscosity, known as viscometry, using a device made up of two concentric cylinders, arranged vertically with height L, with the test substance held between the cylinders (Mallock, 1888, 1896; Couette, 1890). The inner cylinder is locked in a stationary position and the outer cylinder is rotated as indicated in Figure 6.8. From measurements of the angular velocity of the outer cylinder and the torque on the inner cylinder, the viscosity can be determined, as in Equation 6.27. \ufffd b2\u2212a2 \ufffd \u03bc \u00bc \ufffdTq \u00f06:27\u00de 4\u03c0\u03a9ba2b2L Bilgen and Boulos (1973) give the following equation for the moment coefficient for an annulus with inner cylinder rotation with no axial pressure gradient for laminar flow. 188 Rotating Flow \ufffd \ufffd0:3 Cmc \u00bc 10 b\u2212a Re \u22121 \u00f06:28\u00de ma where the rotational Reynolds number, Re m, is based on the annulus gap, \u03c1\u03a9a\u00f0b\u2212a\u00de Re m \u00bc \u03bc \u00f06:29\u00de for Re m<64 \u00f06:30\u00de The equivalent relationship for the moment coefficient from the linear theory, Equations 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001229_tcst.2010.2055566-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001229_tcst.2010.2055566-Figure5-1.png", + "caption": "Fig. 5. Optical side of the experimental setup.", + "texts": [ + " The experimental setup consists mainly of a MFDM wavefront corrector, a Shack\u2013Hartmann type wavefront sensor used to measure the shape of the wavefront reflected off the MFDM, a control computer running the proposed control algorithms, and a laser diode with 661 nm wavelength used as the reference light. The wavefront corrector is a prototype MFDM consisting of a 60 mm diameter, 1 mm thick layer of EFH1 magnetic fluid, and an array of 19 electromagnetic coils used to control the shape of the fluid surface. A snapshot of the optical side of the experimental setup is shown in Fig. 5. In the experimental evaluation, the reference signal is given by where is a static signal chosen to represent a typical static reference wavefront shape signal and is given as m (42) The signal represents the time-varying component of the wavefront aberrations and can be represented as a low frequency sinusoidal signal or a low-pass filtered random signal to be consistent with the low frequency content of the aberations in the eye [24]. The performance of the designed PID controller in tracking static reference signals is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001652_j.jsv.2008.12.018-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001652_j.jsv.2008.12.018-Figure1-1.png", + "caption": "Fig. 1. Diagram of driveshaft.", + "texts": [ + " It is therefore the subject of this paper to present experimental results, derivation of the secondary moment under no torque load, derivation of a complex driveshaft model and subsequent correlation of the simulation results with the experimental results. The driveshaft of consideration is a commercially available driveshaft outfitted on a current production vehicle. The driveshaft is of a two piece design incorporating a CV joint at the transmission end of the shaft and two non-constant velocity joints located in the center and rear axle end of the shaft as pictured in Fig. 1. The CV joint is of the plunging type with six balls allowing for the angular misalignment. The center bearing consists of a ball bearing supported by a diaphragm type flexible rubber bushing. The non-constant velocity joint is located just to the front of the center bearing assembly. The second non-constant velocity joint is located at the rear axle end of the driveshaft. In addition, a vibration absorber, internal tuned damper (ITD), is incorporated in the front shaft to minimize vibrations due to the first bending mode of the driveshaft" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003469_j.oceaneng.2013.01.001-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003469_j.oceaneng.2013.01.001-Figure2-1.png", + "caption": "Fig. 2. Test-bed of CR200.", + "texts": [ + " In Section 2, the dynamic equations for subsea legged robot are derived by using joint variables, body coordinates, hydrodynamic forces, and interactions with seabed. The Joint torque constraint equation is derived as an inequality form through a frictional constraint condition in Section 3. The mathematical framework for analysis of mobility and agility are described in Section 4, and the simulation results for verification of the proposed method are given in Section 5 with example studies of 4 and 6-legged robot. In this section, the dynamic equation of multi-legged robot such as shown in Fig. 2 is derived to solve the acceleration of the robot\u2019s body center, which is based on the mobility and agility analysis of seabed walking robot. This equation is described as the relationship between the joint torque and the acceleration of body center. Multi-legged robot is composed of several legs and rigid body supported by the legs. The dynamic equations of the body is described as Q \u00bc IB \u20acuB\u00feQ B\u00feDB, \u00f01\u00de where Q \u00bc f x f y f z Zx Zy Zz h iT is the force (f) and moment (Z) at center of body, IB is the inertia term, and \u20acuB \u00bc \u20acrx \u20acry \u20acrz _ox _oy _oz h iT is the term including transla- tional acceleration (\u20acr) and rotational acceleration ( _o) at the center of body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001095_s11044-010-9236-5-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001095_s11044-010-9236-5-Figure8-1.png", + "caption": "Fig. 8 Rigid-flexible double pendulum", + "texts": [ + " Since new constraints that need to be implemented for the absolute nodal coordinate formulation are the orthonormality conditions given by CO(eik) = 0 and the coordinate mapping equations CR(eik,pik) = 0 which are both of simple form as shown by (28) and (32), integrations of the absolute nodal coordinate formulation to existing general-purpose multibody dynamics codes can be achieved in a straightforward manner and the computer implementation can be easier than that of the existing joint coordinate system approach [5]. In this section, several numerical examples are presented in order to demonstrate the use of the proposed joint constraint formulation developed in this investigation. Deformation within the elastic range is assumed in the following examples. 6.1 Rigid-flexible double pendulum The first example consists of a rigid body which is pinned to the ground and connected to another very flexible body by a revolute joint. Figure 8 shows the model of the rigidflexible double pendulum. In this example, the second body is assumed to be very flexible and is modeled using the absolute nodal coordinate formulation. The axis of the first joint is selected to be along the global Y -axis, while the second joint axis makes 30 degrees with the X-axis to obtain a general three-dimensional motion as shown in Fig. 8. This example is given in the literature [5] with the joint coordinate approach, and numerical results obtained using this approach and the one proposed are compared. For this reason, the flexible link is divided into four elements as is given in the literature [5]. The dimensions and the material properties of both links used in this example are given in Table 1 [5]. Figure 9 shows motion of the rigid-flexible double pendulum. The global positions of the first and second links are shown in Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure10-1.png", + "caption": "Fig. 10 Balanced crank-slider mechanism synthesized from the basic CRCM configuration", + "texts": [], + "surrounding_texts": [ + "a o d d f t b\nr o l m\n3\nm p t q\np p o n n p b n\nm s b 7 C\nn l n C s t\nd i s A t\nF n\nJ\nDownloaded Fr\nparallel transmission k1=0 and does not rotate for any motion f the mechanism. m2\nis then used for the moment balance of both egrees of freedom, as in Fig. 4. The result is that the term I1 rops from Eqs. 26 , 28 , 29 , 31 , and 32 . It is evident that or a low inertia, the countermasses should not rotate with respect o the base, besides what is necessary to maintain the moment alance.\nLimitations of the CRCM principle can be the transmission atios that have to remain relatively small in practice, a gear ratio f 8 is already high , and the shape of a CRCM that may become arge to obtain the specific amount of inertia and mass. These and\nore limitations are studied in Ref. 12 .\nCRCM-Balanced 2DOF Parallel Mechanisms In this section, three new CRCM-balanced 2DOF parallel echanisms are synthesized by using the CRCM-balanced double\nendulum of Fig. 1. It is shown that the balancing conditions and he inertia equations for these parallel mechanisms can be derived uickly.\nAny combination of one or more balanced double or single endula results into a balanced mechanism. Therefore, two double endula balanced, as in Fig. 1, can be combined, such that their rigin is at the same location and they form the parallel mechaism of Fig. 6. The end point of each double pendulum does not eed to coincide but can be anywhere as long as the links remain arallel. The mass at the end point and its inertia can be balanced y both links or by only one of them. The other then is still ecessary to balance the mass and inertia of the link itself.\nSince the angular velocities of parallel links are equal for the oment balance of two parallel links, only one CRCM is necesary. This means that there are only two CRCMs necessary, which oth can be constructed compactly near the base, as shown in Fig. . This is a configuration described in Ref. 14 . The former RCMs become fixed countermasses. It is also possible to derive this parallel mechanism by combiation of an idler loop 15 and a CRCM-balanced double penduum, as shown in Fig. 8. Also in this case, only two CRCMs are ecessary, and by using the countermass of the idler loop as a RCM, they can be constructed near the base. This result is hown in Fig. 9 and it has only one fixed countermass instead of wo, as in the configuration of Fig. 7.\nWith the equations of the angular momentum of the balanced ouble pendulum being known, the inertias of the CRCMs and the nertia equations of the mechanism can be calculated quickly by imply adding the equations of each individual double pendulum. s an example, the angular momentum and inertia equations of\nA\nO\nig. 6 2DOF balanced parallel mechanism obtained by combiation of two CRCM-balanced double pendula\nhe parallel mechanism of Fig. 7 are calculated. Therefore, it is\nournal of Mechanical Design\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201\nassumed that half of the mass m2 and inertia I2 at the end point is balanced by each double pendulum. By changing the notations slightly with m2 =m2,1 , m1 =m1,1 , I2 = I2,1 , and I1 = I1,1 in which the additional index 1 represents double pendulum 1, the angular mo-\nNOVEMBER 2009, Vol. 131 / 111003-5\n6 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "i\nC r\nA r f\n4\nu f t d\nd b C c s t a n\nfi w t t\n1\nDownloaded Fr\n+ I2\n2 +\nm2\n2 l1 2 + m1,2 l1 2 + I1,2 \u03071 44\nn which links 1 and 2 have lengths l2 and l1, respectively. \u03071 and\n\u02d9 2 depend on \u03071 and \u03072 as\n\u03071 = \u2212 \u03072 45\n\u03072 = \u03071 + \u03072 46\nombining 1hO,z and 2hO,z and substituting Eqs. 45 and 46 esult into one equation for the angular momentum:\n1+2 hO,z = I2 + I2,1 + I1,2 + m2 + m1,2 l2 2 + m2,1 + m2,2 l2 2\n+ m2 + m2,1 l1 2 + m1,1 + m1,2 l1 2 + k1I1,1 + k2I2,2 \u03071\nI2 2 + I2,1 + m2 + m1,2 l2 2 + m2,1 + m2,2 l2 2 + k2I2,2 \u03072 47\ns before, from the angular momentum, the equations for the educed inertias I 1 red and I 2\nred can be derived. The conditions for the orce balance of the mechanism in this case are\nm1,1 = m2\n2 + m2,1\nl1 l1 , m2,1 = m2l2 2l2\nm1,2 =\nm2l1\n2l1\n, m2,2 = m2\n2 + m2,2\nl2\nl2\nCrank-Slider and Four-Bar Mechanisms In this section, the various CRCM-balanced double pendula are sed to derive CRCM-balanced crank-slider mechanisms and our-bar mechanisms, and it is shown that for these mechanisms, he balancing conditions and the inertia equations can also be erived quickly.\nBy restricting the motion of the end point of the balanced ouble pendulum to move along a specific trajectory, CRCMalanced crank-slider mechanisms can be obtained from the RCM-balanced double pendula of Figs. 1 and 3\u20135. For the three onfigurations, the CRCM-balanced crank-slider mechanisms are hown in Figs. 10\u201312, for which the slider moves along a straight rajectory with offset h. The slider mass then does not rotate, and CRCM is only needed for the moment balance of the link conected to the slider. The advantages of each CRCM configuration remain if the con-\nguration is used as crank-slider mechanisms. Important features, ith respect to many other possible balancing configurations, are\nhat there are no transmission irregularities or singularities, and he mechanism can fully rotate by suitable link lengths .\n11003-6 / Vol. 131, NOVEMBER 2009\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201\nAlso for the crank-slider mechanisms, the conditions for the force and moment balance and the inertia equations can be obtained easily. The procedure is equal to that of the parallel mechanisms: first writing the angular momentum of the double pendulum, which is known, and then substituting the kinematic relations. In these 1DOF crank-slider mechanisms, 2 depends on 1. This relation is easy to find from the second equation of r2 and its derivative:\nr2,y = l1 sin 1 + l2 sin 2 = h 48\nTransactions of the ASME\n6 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "W w\nA a\nS r\nT p s t c\np m s t t l t c s s t\nC a\nF t p\nJ\nDownloaded Fr\nr\u03072,y = l1\u03071 cos 1 + l2\u03072 cos 2 = 0 49\nith \u03072= \u03071+ \u03072, \u03072 can then be written as\n\u03072 = \u2212 l1 cos 1\nl2 cos 2 \u2212 1 \u03071 50\nith\n2 = sin\u22121 h \u2212 l1 sin 1\nl2 51\ns an example, the configuration of Fig. 12 is taken for which the ngular momentum writes from Eq. 37 :\nhO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a \u03071\n+ I2 + m2l2 2 + m2 l2 2 + k2I1,b \u03072 52\nubstituting the kinematic relations of Eqs. 50 and 51 then esults in\nhO,z = I2 + I1,b + m2l2 2 + m1 l1 2 + m2 l2 2 + m2 + m2 l1 2 + k1I1,a\n+ I2 + m2l2 2 + m2 l2 2 + k2I1,b \u2212 l1 cos 1\nl2 cos 2 \u2212 1 \u03071 53\nhe single equation of the reduced inertia is now dependent on the osition of the mechanism. This equation also holds for an uncontrained balanced double pendulum moving along the same trajecory, although then there are two input angles where each has a onstant reduced inertia.\nAnother balanced slider mechanism can be derived from the arallel mechanism of Fig. 7. If the end point of this parallel echanism moves along a straight line through the origin, as hown in Fig. 13, then the two CRCMs can become fixed counermasses. Half of the mass m2 and half of I2 if not a slider is hen balanced by each link that is attached to it. The inertia and ength of these links must be equal and also the links attached to he origin must have equal inertia and length by which they beome moment balanced sets. With these conditions, it is also posible to have the length and inertia of the links attached to the lider to be different from the length and inertia of the links atached to the origin.\nBerestov 10 showed a planar four-bar mechanism balanced by RCMs driven by inner gears. This mechanism can be regarded\nA\nO\nl *\n2\nl *\n1\nl 1\nl *\n1\nl 2\nl *\n2\nm 2\nl 1\nl 2\nig. 13 1DOF crank-slider mechanism without CRCMs obained by restricting the motion of the end point of the 2DOF arallel manipulator\ns a combination of a balanced single and a balanced double pen-\nournal of Mechanical Design\nom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201\ndulum, shown with chain driven CRCMs in Fig. 14. Also for four-bar mechanisms, the different CRCM configurations are applicable just as the substitution of the well-known kinematic relations into the inertia equations of the double and single pendula to obtain the inertia about one of the links. This means that with the equations for the double and single pendula and the kinematic relations the inertia of any four-bar mechanism can be written down easily.\nIt is a special case for which the four-bar linkage becomes a parallelogram. From Fig. 9, and assuming the link between O and A to be fixed with the base, the resulting parallelogram can be balanced, as in Fig. 15, with solely a CRCM. Because the coupler link does not rotate, its center of mass can be located arbitrarily.\n5 3DOF Parallel Mechanisms CRCM-balanced 3DOF planar and spatial mechanisms can be synthesized by combining the CRCM-balanced double pendula. Two examples are the planar 3-RRR parallel mechanism of Fig. 16, which has one rotational and two translational DOFs, and the spatial 3-RRR parallel mechanism of Fig. 17, which has two rotational and one translational DOFs. As described in Ref. 16 , the platforms of these mechanisms can be modeled by lumped masses at their joints, maintaining its original mass, its location of the center of mass, and its inertia tensor. This allows each leg to be balanced individually for which their combination is balanced too. The dimensions of each leg can be different, as long as each leg is balanced.\nTo obtain the inertia equations of these mechanisms, the kinematic relations can be substituted into the angular momentum equations of the double pendula. Since there are multiple closed\nNOVEMBER 2009, Vol. 131 / 111003-7\n6 Terms of Use: http://www.asme.org/about-asme/terms-of-use" + ] + }, + { + "image_filename": "designv11_3_0003601_isie.2011.5984271-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003601_isie.2011.5984271-Figure1-1.png", + "caption": "Figure 1. Main parts and motion idea of the TRAS laboratory system.", + "texts": [ + " The ADRC basic idea, its key elements, and some tuning guidelines are presented in third part. Section IV describes conducted experiments, their goals, made assumptions, and obtained results. Paper ends with few concluding remarks in Section V. TRAS is a laboratory setup designed for educational purposes. System is often used by researchers as a benchmark tool to explore, implement, and evaluate different control methodologies. From a mechanical point of view TRAS has two rotors (front and rear rotor) placed at both ends of a beam (see Fig.1). Rotors are driven by two direct current (DC) motors. The beam is pivoted on a rigid stand. The above construction allows the beam to move freely both in vertical and horizontal plane. Additionally, a movable counterbalance attached to the beam at the pivot point can be used to set a desired equilibrium position. An exemplary mathematical model of TRAS, based on [9], is presented on Fig.2, where subscripts \u201ev\u201d and \u201eh\u201d represent the elements associated with vertical and horizontal motion of the beam respectively, and \u03b1[rad] \u2013 angular position of the beam (output signal), U [V ] \u2013 DC motor input voltage (input/control signal), H\u2217 \u2013 DC motor dynamics, F \u2217 \u2013 function describing transformation of rotational speed of the propeller (\u03c9[rad/s]) into aerodynamic force (F [N ]), L[m] \u2013 lenght of 978-1-4244-9312-8/11/$26" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000891_elan.200904644-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000891_elan.200904644-Figure7-1.png", + "caption": "Fig. 7. Cyclic voltammograms of LDH/GC electrode in 0.01 M pH 7.0 PBS containing 0.1 M KCl in (a) the absence and the presence of (b) 0.06 mM, (c) 0.12 mM, and (d) 0.24 mM H2O2. Scan rate, 100 mV s 1.", + "texts": [ + " However, it was found that the peak current gradually decreased and the peak shape was more unsymmetrical (redox current), which might be due to LDHs partly dissolved and the structure of LDHs somewhat destroyed in the acidic solution. As is well-know, sensitive and accurate determination of a trace of hydrogen peroxide is of great importance, because H2O2 is not only a by-product of several highly selective oxidases, but also an essential mediator in biology, medicine, industry and many other fields [42]. The electrocatalytic reduction of H2O2 by LDH/GC electrode is shown in Figure 7. When H2O2 was added in buffer solution, an increase in the reduction peak was observed at about\u00fe0.176 V with the decrease of the oxidation peak of the Mg-Al-HCF LDHs. No electrochemical response was observed when the cyclic voltammetric scan was carried out at bare the GC or the MgAl-CO3 modified GC electrode. The mechanism of electro- 2129Electrocatalytic Reduction of H2O2 2009 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim www.electroanalysis.wiley-vch.de Electroanalysis 2009, 21, No. 19, 2125 \u2013 2132 catalytic reduction of H2O2 by Mg-Al-HCF LDHs in pH 7" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000158_s0022-0728(80)80276-2-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000158_s0022-0728(80)80276-2-Figure2-1.png", + "caption": "Fig. 2. Cyclic v o l t a m m o g r a m showing oxide film fo rma t ion on a gold e lec t rode surface in a so lu t ion of a) suppor t ing e lect rolyte solut ion conta in ing NaC104, o. i M, p h o s p h a t e buffer, o.o2 M, a t p i t 7 ; b) above suppor t i ng e lec t ro ly te w i th 4 ,4 ' -b ipyr idy l , o .o i M ; c) as in Figure 2b excep t c y t o c h r o m e c, 5 mg /cmK", + "texts": [ + " The electrode reaction, which is not observed in the absence of 4,4'-bipyridyl, corresponds to the reduction and oxidation of the cytochrolne heine iron, the reduction of 4,4 ' -bipyridyl occurring at much lower potentials, about -0.8 V vs. N.H.E. Controlled potential reduction and subsequent reoxidation, followed spectrophotometrically, confirms that the electrode process is due to the heine iron. EGect of 4,4'-bipyridyl and cytochrome c on oxide film formation at the gold surface The cyclic vol tammogram obtained at a gold electrode in the supporting electrolyte, o.i M NaCI04, o.o2 M phosphate buffer p i t 7, over the potential range 0 to !.5 V vs. s.c.e, shows formation of an oxide film on the gold surface [I2~ (Fig. 2a). Addition of 4,4 ' -bipyridyl to the soiution considerably reduces the degree of oxide film formation (Fig. 2b) and this is consistent with adsorption of 4,4 ' -bipyridyl onto the gold surface. Addition of cytochrome c further reduces the degree of oxide film formation (Fig. 2c), though is without effect in the absence of 4,4'-bipyridyl. This suggests some interaction between cytochrome c and the adsorbed bipyridyl resulting in a more stable surface layer. Addition of 4,4 '-bipyridyl is also found to increase the overpotential for oxygen reduction at gold and the overpotential is further increased upon addition of cytochrome c. Again is consistent with adsorption of bipyridyl onto the electrode surface and stabilization of this surface layer by interaction with the cytochrome" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003312_j.jsv.2013.05.022-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003312_j.jsv.2013.05.022-Figure2-1.png", + "caption": "Fig. 2. Rotor\u2013stator structure and magnetic forces on magnet/tooth.", + "texts": [ + " To facilitate the analysis conducted here, this work makes the following assumptions and simplifications: (1) Since one major concern is the effect of the magnet/slot number on the magnetic forces, for simplicity, the forces are equivalently substituted by concentrated loads on each magnet/slot center during theoretical analysis, and the equivalent couple of forces is omitted. This assumption is of fundamental importance, and it has no influence on harmonic orders instead of magnitudes, which contributes to the clarification of the expected relationship. (2) The foundation of this work is the mechanical and magnetic symmetries even when inevitable machining, installation and magnetization errors are included. Note that the errors to-a-certain-degree destroy the symmetries, but symmetries can still be hold because the motor's periodicity always remains. Fig. 2 illustrates a schematic of a rotor\u2013stator structure and magnetic forces on rotor and stator sides. The magnet and tooth numbers are Nm and Ns, respectively. fok; xk; ykg are rotor-fixed (k\u00bc 1) or stator-fixed (k\u00bc 2) coordinates. Without any loss of generality, the horizontal axis xk directs toward the center of the first magnet/tooth. The angle between the first magnet/tooth and the horizontal axis is \u03c81 \u00bc 0, and thus those of the ith magnet and jth tooth are \u03c8 i and \u03c8 j, respectively. Fmti and Fmri denote tangential and radial forces on the ith magnet, respectively; Fstj and Fsrj are the corresponding tangential and radial forces on the jth tooth, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.20-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.20-1.png", + "caption": "Fig. 2.20 Pierre Le Roy\u2019s detent escapement (Wikipedia 2004f)", + "texts": [ + " Two hundred and seventy years later, in 2008, as a tribute to the Grasshopper escapement, the Corpus Clock was built at Corpus Christi College, Cambridge University, in Cambridge, England. It is shown in Fig. 2.19 (Wikipedia 2008). The ever moving grasshopper exemplifies that the time is another dimension of the universe. The spring detent escapement, most commonly used on some nineteenth century\u2019s precision watches, is a type of detached escapement. The early form was invented by the French watchmaker Pierre Le Roy (1717\u20131785) in 1748 (Fig. 2.20), who created a pivoted detent type of escapement (Wikipedia 2004f). It was then generalized in 1783 by the English watchmaker Thomas Earnshaw (1749\u20131829) (Wikipedia 2005) with his standard spring detent escapement and used until mechanical chronometers became antediluvian (Fig. 2.21). Although John Arnold (1736\u20131799) and Swiss watchmaker Ferdinand Berthoud (1727\u20131807) both had their own design in 1779, neither of their designs could match Earnshaw\u2019s design in popularity. Due to the virtual absence of sliding friction between the escape tooth and the pallet during impulse, the spring detent escapement could be made more accurate than lever escapements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002442_978-0-387-92904-0_2-Figure2.10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002442_978-0-387-92904-0_2-Figure2.10-1.png", + "caption": "Fig. 2.10 The ascending fabrication platform proposed by [47]", + "texts": [ + " Create a solid or surface model on a CAD system 2. Export the CAD model 3. Add support structures 4. Specify the build style variables and parameters necessary for slicing 5. Slice the computer model to generate the information that controls the SL apparatus 6. Build the model using the slice file 7. Post-process and clean the part 8. Post-curing to complete the cure process. The block diagram of the stereolithography system as proposed by Hull [47, 48] is shown in Fig. 2.9. Hull also proposed other stereolithographic strategies as shown in Fig. 2.10. In this system the physical object is pulled up from the liquid resin, rather than down and further into the liquid photopolymeric system [47]. The radiation passes through a UV transparent window. In order to minimize the amount photopolymerisable material required for the fabrication process, Murphy et al. [52] proposed a stereolithographic method and apparatus in which a membrane separates two liquid phases. The system (Fig. 2.11) comprises a nonpolymerisable fluid phase, an impermeable movable membrane positioned on top of the fluid phase, a photopolymerisable liquid resin positioned on top of the membrane and a radiation source positioned above the polymerisable material [52]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002314_icsse.2011.5961870-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002314_icsse.2011.5961870-Figure5-1.png", + "caption": "Figure 5. Gear fault: chipped tooth (a), gear tooth broken (b), gear crack (c)", + "texts": [ + " The structural faults such as unbalance, misalignment and looseness were analyzed. The five types of gear faults are used in experiment with gear tooth broken, chipped tooth, gear crack, gear tooth broken combine with chipped tooth and gear crack combine with chipped tooth, respectively. The misalignment was simulated by adjusting the high of the gearbox base plane using thin shims. Adding mass on the output of the gear shaft leads to mechanical unbalance. The detail descriptions of the faults are shown in Table 1. The gear failures are shows in Fig. 5. The vibration signal was acquired by four accelerometers located on the outer case of the gearbox. The accelerometers were used to measure the gearbox vibration signals from horizontal and vertical direction. In the axial directions, vibration signals were ignored since test rig used the spur gear in which axial direction vibration is not obvious. The shaft speed 1184 rpm was obtained by one laser speedometer. Forty ICSSE 2011 continuous measurements were recorded for each condition. The recorded time length of each time is 2 seconds, the number of data is 16,384 and sampling rate is 8,192 Hz" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure4-1.png", + "caption": "Fig. 4 Configuration 1 of the example", + "texts": [ + " 2 The configuration when a five-bar linkage configuration hanges to a four-bar linkage configuration 74501-2 / Vol. 130, JULY 2008 om: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash 3 Examples of Planar Mechanisms Figure 3 shows a planar mechanism with five links. The original adjacency matrix of the mechanism, A0, is the one in Eq. 1 with Fm=2. There are five more possible configurations of the mechanism as there are five 1 among upper triangle of A0. Each one will result in a different adjacency matrix as shown below. 1 If Slider 2 is fixed to Link 1, as shown in Fig. 4 a , the prismatic pair between Link 1 and Slider 2 is frozen. The mechanism is termed as Configuration 1 and is shown in Fig. 4 b . It works as a general four-bar linkage mechanism with DOF F=1. The adjacency matrix A1 can be obtained by multiplying \u22121 on A0 1,2 and A0 2,1 , i.e., A1 = 0 \u2212 1 0 0 1 \u2212 1 0 1 0 0 0 1 0 1 0 0 0 1 0 1 1 0 0 1 0 8 2 If Slider 2 is attached together with Link 3 by a pin, P, as shown in Fig. 5 a , the revolute pair between Slider 2 and Link 3 is then frozen. The mechanism is termed as Configuration 2 and is shown in Fig. 5 b . Its adjacency matrix, A2, can be obtained by multiplying \u22121 on A0 2,3 and A0 2,3 , i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003596_ccgrid.2012.144-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003596_ccgrid.2012.144-Figure3-1.png", + "caption": "Figure 3. Gravity acceleration vector decomposition.", + "texts": [ + " The human body can be considered as a system composed of several joint links, whose mutual relative position represents the primary information for detecting the corresponding body posture. Such information can be obtained by positioning accelerometer sensors in specific part of the body, in order to get sensor orientation (angular position) in relation to the gravity acceleration vector. The relationship between the orientation of a rigid body, i.e. the accelerometer sensor, and the g vector with respect to the reference coordinates Ox0y0z0, is depicted in Fig. 3. In static conditions, an accelerometer is able to returns proper values through which it is possible to calculate the angle \u03b8 between the gravity vector, having a vertical direction, and every axis of the sensor. In particular, taking into consideration the accelerometer subject to static acceleration, the variation of the component of g on a single generic axis and the corresponding value of the axis angle with respect to the g vector are shown in Table I. The architecture of the proposed posture recognizing system, depicted in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001719_tcst.2008.2004428-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001719_tcst.2008.2004428-Figure1-1.png", + "caption": "Fig. 1. Configuration of ETS-VI.", + "texts": [ + " Moreover, from the maximum value and the minimum value , the convex decomposition of the vector is given as (31) Then, the coefficient matrices of the closed-loop system (17) can be described as follows: (32) with (33) By using the relationship to the bounded real lemma (19), we can obtain a stability condition (34) where the matrices and are defined as (35) Therefore, LPV DDFV and DVDFB controllers can be designed similarly to the LTI optimal controllers by applying the same schemes to the simultaneous matrix inequalities (36) We now apply our design methods to the flexible spacecraft attitude control problem of the ETS-VI satellite [15], [16] and confirm the capability of the proposed symmetric controllers described inSections III, IV, and V using a mathematical model of ETS-VI. A view of ETS-VI is shown in Fig. 1. ETS-VI is a three-axis stabilized geosynchronous spacecraft launched in 1994 for advanced communication experiments. It has a pair of large lightweight solar paddles which rotate around the pitch axis at the orbital rate. Controlling such a large flexible spacecraft with high accuracy requires control of its structural vibrations. Many types of robust controllers including control have been designed for this problem and their performances evaluated on orbit [16], and follow-on studies on synthesis and gain-scheduled control have also been conducted [17], [18]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003548_apec.2012.6166044-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003548_apec.2012.6166044-Figure11-1.png", + "caption": "Fig. 11. Experimental setup of five-phase PM machine drive", + "texts": [ + " The objective function for calculation of optimal currents can be modified to decrease high peak currents and higher order harmonics in the optimal currents at the expense of increased ohmic losses and/or torque pulsations. This subject is out of scope of this paper. In this section experimental results on the PM machine of the previous section are presented to verify the proposed FTC strategy. Fig. 10 shows block diagram of the proposed control algorithm to supply the motor with optimal currents. A hysteresis current controller is used to generate gate drive signals for the voltage source inverter. DS1104 dSpace controller is used to implement the control scheme. Fig. 11 shows experimental test setup. A five-phase IGBTbased voltage source inverter is used to supply the motor. The information of the stator currents and the rotor position are obtained using Hall Effect current sensors and a rotary optical encoder, respectively. Steady state tests are conducted with speed reference of 1000 rpm and load torque of 1.15 N.m. Output torque of PM machine under different machine faults are shown in Fig. 12. Fig. 13 shows output torque of the PM machine for healthy condition and different line faults" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003025_978-3-642-33509-9_38-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003025_978-3-642-33509-9_38-Figure2-1.png", + "caption": "Fig. 2. Schematic representation of the Qball-X4", + "texts": [ + " The generated thrust Ti of the ith motor is related to the ith PWM input ui by a first-order linear transfer function Ti = K \u03c9 s+ \u03c9 ui (1) where i = 1 2 3 4 and K is a positive gain and \u03c9 is the motor bandwidth. K and \u03c9 are theoretically the same for the four motors but this may not be the case in practice. It should be noted that ui = 0 corresponds to zero thrust and ui = 0.05 corresponds to the maximal thrust that can be generated by the ith motor. Geometry\u2014A schematic representation of the Qball-X4 is given in \u201cFig. 2\u201d. The motors and propellers are configured in such a way that the back and front (1 and 2) motors spin clockwise and the left and right (3 and 4) spin counterclockwise. Each motor is located at a distance L from the center of mass o and when spinning, a motor produces a torque \u03c4i which is in the opposite direction of that of the motor as shown in \u201cFig. 2\u201d. The relations between the lift/torques and the thrusts are uz = T1+T2+T3+T4; u\u03b8 = L(T1\u2212T2); u\u03c6 = L(T3\u2212T4); u\u03c8 = \u03c41+\u03c42+\u03c43+\u03c43 (2) The torque \u03c4i produced by the ith motor is directly related to the thrust Ti via the relation of \u03c4i = K\u03c8Ti with K\u03c8 being a constant. In addition, by setting Ti = Kui from (1), the relations (2) can be written in a compact matrix form as \u23a1 \u23a2\u23a2\u23a3 uz u\u03b8 u\u03c6 u\u03c8 \u23a4 \u23a5\u23a5\u23a6 = \u23a1 \u23a2\u23a2\u23a3 K K K K KL \u2212KL 0 0 0 0 KL \u2212KL KK\u03c8 KK\u03c8 \u2212KK\u03c8 \u2212KK\u03c8 \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a2\u23a2\u23a3 u1 u2 u3 u4 \u23a4 \u23a5\u23a5\u23a6 (3) where uz is the total lift generated by the four propellers and applied to the quadrotor UAV in the z-direction (body-fixed frame). u\u03b8, u\u03c6, and u\u03c8 are respectively the applied torques in \u03b8, \u03c6, and \u03c8 directions as illustrated in \u201cFig. 2\u201d. L is the distance from the center of mass to each motor. Dynamics of the Quadrotor\u2014A commonly employed quadrotor UAV model is mx\u0308 = uz(cos\u03c6 sin \u03b8 cos\u03c8 + sin\u03c6 sin\u03c8) my\u0308 = uz(cos\u03c6 sin \u03b8 sin\u03c8 + sin\u03c6 cos\u03c8) mz\u0308 = uz(cos\u03c6 cos \u03b8)\u2212mg (4) where x, y and z are the coordinates of the quadrotor UAV center of mass in the earth-fixed frame. \u03b8, \u03c6, and \u03c8 are the pitch, roll and yaw Euler angles respectively. m is the mass. If we fix the yaw angle to zero (\u03c8 = 0) and consider the roll and pitch angles to be very small then a simplified linear model can be obtained in hovering conditions [11]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002229_tpas.1971.292930-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002229_tpas.1971.292930-Figure1-1.png", + "caption": "Fig. 1. Shell of finite thickness and infinite length current filament inside shell.", + "texts": [ + " The expressions derived for the magnitudes mentioned above correspond to both proximity effect and skin effect. ASSUMPTIONS The same assumptions are valid as in Part I. Paper 70TP 600 PWR, recommended and approved by the Switchgear Committee of the IEEE Power Group for presentation at the IEEE Summer Power Meeting and EHV Conference, Los Angeles, Calif., July 12-17, 1970. Manuscript submitted February 17, 1970; made available for printing May 26, 1970. Development of Equations The current filament is placed at a distance b from the axis of the shell, Fig. 1. The eddy current density distribution in the material of the shell i(r,4) is given by Eq. 1 Part I while the magnetic vector potential within the material Ao(r,k) is given by Eq. 2, Part I. These two equations have as solutions respectively: L&(,w) = ZL [F 1, ( jI/VvY)+KC K, (j(Zpr)] co$st Ao(sw)=( [V L, (AI)r CK44V'tpr cosVvp (1) (2) Both modified Bessel functions In and Kn appear now. The magnetic vector potential due to the current in the filament is given by Eq. 13, Part I, or A t (pin9) =Mzo", + "lte-eo)}laV/o = I cIC-(VZC ) -L,oXictLI (\"I' a)( ~~~~I0 2 x 8) Since only the z-component of the magnetic vector potential exists, the following equations can be written, connecting flux density components and magnetic vector potential (21) Is\u00b0 W(3 Pa) 2-j 117 4 \\1: (avi-t _ __ (1- 1%d1 'a) (I0 Ct-c (I The total magnetic vector potential inside the shell will be: A; (p,) = AI (eO) + As>, (?,e) (1 1) Outside the shell, the magnetic vector potential will be due to the distribution of eddy currents in the shell only. For the current in an infinitesimal element rdrdo of the crossection it is: d- Aa,e (?,, ) =u[o (i-) q) tnRv rdxciD&~~~2 (12) where i(r,O) is given by Eq. 1 and R1 is the distance between points P1(r,4) and P3(p,O) Fig. 1. V\\=. -((-L$(coosvnc. co05 v + (22) (9) By using Eqs. 2, 1 1 and 14 as well as 21 and 22, the four Eqs. 17, 18, 19 and 20 representing the boundary conditions become: .~>'1L~o'S .Ve -- II _ 1122. 112 l 5L X%Cs,_;Pl ~~~~~ -~~~~~ 3~~I2.~c .W2-Pcv - s (jLpv)FVI y 112 C( i 4-)C + atL Lv%4, CLce~ 2.8 M,=j{P I r -L Ta(i PC) F + rth( lp) j'pc j\" 3pc -1W(ilpc) Fv, lc m Aj-SAv,i (p itv\\ ) with p>r Integrating Eq. 8 over the whole crossection area of the shell, with respect to r from a to c and with respect to k from 0 to 2ir yields A6,e(" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002238_elan.201100540-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002238_elan.201100540-Figure2-1.png", + "caption": "Fig. 2. Scheme of the screen-printed electrode adopted in this work.", + "texts": [ + " In fact in order to avoid the contact with the powder, the operators were dressed with lab dresses, gloves, mask and glass. Also a dedicated fume hood was used during sample preparation and analysis. T O P IC A L C L U S T E R Electroanalysis 2012, 24, No. 3, 581 \u2013 590 2012 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim www.electroanalysis.wiley-vch.de 583 In order to make the production of the prototype easier, the use of an electrochemical sensor composed only of two electrodes (working and reference) is highly suggested. Thus, a screen-printed electrode with these characteristics was produced (Figure 2). In this case, the working surface area (0.007 cm2) was lower than that of the biosensor with a three electrode configuration (0.03 cm2), thus using the enzymatic membrane previously optimised [1,25] very low current values were obtained. In order to observe a detectable current value with satisfactory accuracy, the enzymatic units should be increased; however this choice was avoided taking into consideration that the inhibition of BChE by nerve agent is irreversible and thus increasing the enzyme amount, the sensitivity of the biosensor towards the nerve agent decreases [1]", + " In this work the storage stability of the biosensor developed was monitored measuring the degree of inhibition vs. 60 ppb of paraoxon using the \u201cthree step\u201d measurement. As storage conditions, the biosensor was maintained dry at room temperature. As shown in the Figure 6a, the enzymatic activity decreases after about 1 month; however, up to the period tested, the biosensor retained the original sensitivity for paraoxon detection (Figure 6 b). After the development and the analytical characterization of BChE biosensor based on screen-printed electrode made only with the reference and working electrodes (Figure 2), the biosensor was integrated in the prototype (Figure 7 a). The prototype was developed in conjunction with Strictes+AeG Company and composed of a cell described in detail in Experimental (Section 2.2 and showed in Figure 1). The biosensor was inserted in the cell and the substrate solution was added by using a syringe (Figure 1 b). The system was able to detect the nerve agent in gas phase because the device is composed of a little ventilator/fan able to sampling 20\u201325 L of air/min, and a canal that connects the fan with the electrochemical cell, in order to gurgle the air sampled by the fun into the electrochemical cell, thus in the solution in which the biosensor is in contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001212_1.2991136-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001212_1.2991136-Figure1-1.png", + "caption": "Fig. 1 Gear model", + "texts": [ + " Manuscript received December 3, 007; final manuscript received May 7, 2008; published online October 8, 2008. eview conducted by Avinash Singh. Paper presented at the ASME 2007 Design ngineering Technical Conferences and Computers and Information in Engineering onference DETC2007 , Las Vegas, NV, September 4\u20137, 2007. ournal of Mechanical Design Copyright \u00a9 20 om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201 2 Equations of Motion A simplified purely torsional model of a gear pair supported by rigid mounts including time-varying mesh stiffness and backlash is considered Fig. 1 . Gear teeth are supposed to be errorless and unmodified so that the relative normal displacement on the base plane with respect to rigid-body positions reads a list of symbols is given in the Nomenclature , = cos b Rb1 1 + Rb2 2 1 The elastic component of the mesh force can be expressed as FM t = km 1 + t if 0 normal conditions 2a FM t = 0 if \u2212 J 0 loss of contact 2b FM t = km 1 + t + J if \u2212 J back strike 2c Because of engine speed fluctuations, rigid-body angular velocities 1 pinion and 2 gear are not constant and generate inertial terms proportional to angular accelerations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000859_tmech.2009.2032180-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000859_tmech.2009.2032180-Figure2-1.png", + "caption": "Fig. 2. Model of HRP2 humanoid robot.", + "texts": [ + " [31] proposed a method for planning a foot-step without considering the upper-body motion. Stilman et al. [32] proposed the manipulation planning of a movable object. Harada et al. [33] proposed the pushing manipulation of a heavy object placed on the ground. The contribution of this research is to combine the walking pattern generator and the collision-free motion planner. For this purpose, we newly supply a two-stage time-parametrized framework for the whole-body motion planning of a humanoid robot walking on flat/rough terrain while maintaining the dynamic balance. Fig. 2 shows a model of the humanoid robot used in this research. Let p\u2217/\u03c6\u2217 be the 3-D vectors of the position/orientation of the coordinate frame fixed to a part of the robot. The subscripts Fj, Hj, B, and G denote the jth foot, jth hand, waist, and CoG, respectively. In addition, let P and L be the linear and angular momentum about the CoG of the robot, respectively, and g = [0 0 g]T be the gravity force vector. We assume that the 3-D models of the robot and the environment are known. These models are used for collision checking" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000031_6.2008-7413-Figure16-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000031_6.2008-7413-Figure16-1.png", + "caption": "Figure 16. 3-D simulation of grounded JAviator (left), the JAviator Control Terminal (middle), and 3-D simulation of airborne JAviator (right).", + "texts": [ + " In response to this packet, the GCS generates a packet containing new navigation data and sends it back to the FCS. The sensor data packet sent by the FCS contains roll, pitch, yaw, and altitude values and their first derivatives as well as x- and y-position data and their second derivatives. Upon receiving a sensor data packet, the GCS responds with a navigation data packet containing new values for roll, pitch, yaw, and altitude. Both sensor and navigation data use the same packet format, as described previously and depicted in Figure 14. The Control Terminal, presented in Figure 16 (middle), displays the sensor data received from the FCS graphically and provides a means for piloting the connected Plant, either the real JAviator or the MockJAviator. The main functionality lies in monitoring and modifying the Plant\u2019s attitude and altitude. This is accomplished with four meters that indicate the \u201ccurrent\u201d and \u201cdesired\u201d roll, pitch, yaw, and altitude values through red and green needles, respectively. In case of the altitude meter, it is also possible to display and set the thrust directly without involving an altitude controller", + " In order to allow for studying flight behavior, the Control Terminal supports recording and replaying of arbitrary flights. Any recorded flight can thus be visualized while replaying. Moreover, the 3-D environment provides the opportunity to view a flight from different perspectives in the 3-D scene, freely adjustable via the rotation, scaling, and translation settings. There are also activatable helper planes that describe the desired attitude and altitude of the JAviator. These planes are activated in both Figure 16 (left) and Figure 16 (right) depicting the grounded and airborne JAviator, respectively. In this sense, the 3-D representation extends the navigation environment by offering full 3-D motion tracking and close-to-reality JAviator visualization effects. In order to study and improve flight stability and system performance, we have extended the GCS with a logging system that enables real-time tracing during flights. This logging system makes use of IBM\u2019s TuningFork,25 an Eclipse-based performance analysis and visualization tool for real-time applications with support for Java, JVM, C++, and Linux" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002650_s12540-011-0223-z-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002650_s12540-011-0223-z-Figure1-1.png", + "caption": "Fig. 1. Schematic drawing of Direct Laser Melting system.", + "texts": [ + " One experiment used a heat source (a fiber laser beam) on a steel die substrate without metal powder to analyze the surface melted area. The other exper- iment used laser melting with metal powder (Fe-Cr and FeNi powders) for an analysis of the deposited layer on the steel die substrate by single-track melting with different laser processing parameters. In the laser beam melting experiments, a fiber laser beam with a diameter of 0.08 mm was used. The laser-melted layer was fabricated with a laser power of 50 W to 200 W and a scan rate of 3.66 mm/s to 366.22 mm/s. Figure 1 shows an illustration of the apparatus used for the laser surface treatment. The radiation source is a YLR-200 CW Ytterbium fiber laser manufactured by IPG Photonics with a maximum power P of 200 W, a wavelength \u03bb of 1075 nm and a laser beam diameter d of 0.08 mm in the focal position. A scanner by Scanlab (hurrySCAN\u00ae20) was used to control the laser scanning method. To produce an oxide-free coating, the chamber was shielded using argon gas (5 l/min) in all experiments. The vertical movement of the cylinder was driven by a motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003664_jfm.2012.480-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003664_jfm.2012.480-Figure1-1.png", + "caption": "FIGURE 1. (a) Definition sketch in two-dimensional polar view and (b) sample deformation in three-dimensional perspective view. Disturbances \u03b71 and \u03b72 are constrained by the belt support extending over \u03b61 \u2261 cos \u03b81 6 cos \u03b8 6 \u03b62 \u2261 cos \u03b82. Lengths \u03b6 are scaled by R while lengths \u03b7 are left unscaled.", + "texts": [ + " Rayleigh\u2019s predictions have been verified experimentally for immiscible drops by Trinh, Zwern & Wang (1982) and Trinh & Wang (1982) and for free drops in microgravity by Wang, Anilkumar & Lee (1996). For drops with viscosity (Chandrasekhar 1961; Miller & Scriven 1968; Prosperetti 1980b) or moving contact lines (Davis 1980), \u03a6 6= 0, and the growth rates typically have real parts so that disturbances can grow or decay. When decay happens, it can occur in an under-damped or over-damped fashion, of course. In this paper, we study the linear stability of two coupled spherical-cap surfaces made by constraining a spherical drop with a solid support (figure 1). The solid support conforms to the spherical surface and extends between two latitudes, \u03b82 6 \u03b8 6 \u03b81, forming a spherical belt. The resulting free surface consists of two spherical caps (disconnected) which are coupled through the liquid beneath (connected). The interfaces are pinned at the edges of the belt and no-slip conditions, as appropriate to viscous liquids, are enforced along the belt. Integro-differential equations for the free-surface disturbances are derived. These operator equations are then solved, restricting to axisymmetric disturbances", + " The operator eigenvalue equation is reduced to a truncated set of linear algebraic equations using a spectral method on a constrained function space, as described in Part 1. The eigenvalues/modes are then computed from a nonlinear characteristic equation, which depends upon material properties and the size/location of the constraint. We conclude with some remarks on the computational results. Consider an unperturbed spherical droplet of radius R, constrained by a spherical belt over the polar angle \u03b82 6 \u03b8 6 \u03b81 in spherical coordinates (r, \u03b8), as shown in the definition sketch (figure 1). The drop interface is disturbed by time-dependent free-surface perturbations, \u03b71(\u03b8, t) and \u03b72(\u03b8, t), which are assumed to be axisymmetric and small. No domain perturbation is needed for linear problems, thus the domain is the combination of the regions internal to and external to the static droplet: Di \u2261 {(r, \u03b8) | 0< r 6 R, 0 6 \u03b8 6 \u03c0}, (2.1a) De \u2261 {(r, \u03b8) | R< r <\u221e, 0 6 \u03b8 6 \u03c0}. (2.1b) The interface separating the interior and exterior fluids (internal boundary) is defined as the union of two free surfaces and one surface of support: \u2202Df 1 \u2261 {(r, \u03b8) | r = R, \u03b81 6 \u03b8 6 \u03c0}, (2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001287_j.forsciint.2010.10.002-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001287_j.forsciint.2010.10.002-Figure2-1.png", + "caption": "Fig. 2. Photograph of the screen-printed electrodes.", + "texts": [ + " Manual screen-printer (Shengjiang, GuangDong, China) and SC-2000 gas chromatograph (Chongqing Sichuan Instrument Analyzer Company) were employed in this study. 2.3. Preparation of screen-printed electrodes The simple production steps of the electrode are shown schematically in Fig. 1. Firstly, three sliver strips were printed on polyvinyl chloride (PVC) basal lamina. Then carbonic working electrode, carbonic counter electrode and Ag/AgCl reference electrode were completed in order. At last, the surface layer was printed with the insulator. The working area was a disc with a diameter of 3 mm. Fig. 2 is the photograph of the screen-printed electrodes. 2.4. Preparation of the GNPs\u2013MWCNT\u2013Nafion hybrid composites Colloidal gold nanoparticles (GNPs) were prepared according to the procedures mentioned before [26] by adding 2.5 mL 1% (w/w) of sodium citrate solution to 100 mL of boiling aqueous solution containing 1 mL of 1% (w/w) HAuCl4 and stirring until the solution discloses claret-red. Preparations were stored in brown glass bottles at 4 8C. The diameter of the resulting gold nanoparticles measured by scanning electron microscopy was about 20 nm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001249_epepemc.2008.4635602-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001249_epepemc.2008.4635602-Figure9-1.png", + "caption": "Fig. 9. Complex function at two different saturation levels and given voltage test phasor depending on the rotor position.", + "texts": [ + " Description As already mentioned, the INFORM method uses test algorithms by utilizing voltage steps and measuring the current response, either using a special measuring sequence interrupting the PWM control or, as shown here, is integrated into an intelligent current control loop. For detection of the axis with maximum (or minimum) magnetic conductance, a sequence of voltage space phasors uS is applied to the PSM via the inverter and the current reaction diS / d\u03c4 is measured. We define (1) By the reason of saliency effects is a 180\u00b0- periodic function, which can be modelled as a circle in the complex plane using the parameters arg(uS) = \u03b3U and rotor angular position \u03b3. (Fig. 9; details are given in [17]): (2) For measuring explicitely, test voltage space phasors uS are applied to the motor during operation. The respective current changes diS / d\u03c4 are measured. Turning the rotor from POS1 to Position POS3 in Fig. 9, a half circle of with radius \u0394y and offset y0 is produced. Applying a positive flux-parallel current iSd increases the circle radius and can be used to improve the INFORM capability of a drive. Fully symmetric PSMs with almost no saturation yield a very small radius \u0394y. 2278 2008 13th International Power Electronics and Motion Control Conference (EPE-PEMC 2008) Combining measurements in different test directions of uS transforms the information of into an offsetfree circle with (3) B. Initialisation The ability of the flux-parallel current component to change the radius \u0394y can be used to distinguish between the two possible solutions (\u03b31 = \u03b3 , \u03b32 = \u03b3 +180o ) at initialization state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002173_304-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002173_304-Figure2-1.png", + "caption": "Figure 2 . A superficial Figure 3. Contact between smooth sphere gap between two deformed", + "texts": [ + " Assuming that the hemispheres deform elastically and that the stress distribution is given as the square root of the parabolic fLinction of the radial Elastic deformation of rough spheres, cylinders and annuli 1473 distance r , Hertz (1881) obtained expressions for both the radius of contact rH and the displacement (or deformation) of the surface w(r) . They are 3PR 113 Y E = (4) and If the deformed spheres shown in figure 1 are now separated by a distance d and still retain their deformed shape as shown in figure 2, the separation distance of two points on the surface, M and N, at a distance r from the centre line may then be approximated by (8) r 2 2 R z= d+ -+ w(v) - w(0) where w is a displacement function. It is interesting to observe in equation (8) that a parameter d, denoting the minimum separation distance, has been included to take into account the roughness of the spheres. Tt will be shown in the following section that such a parameter will depend on the surface topography and material properties as well as the applied load" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure2-1.png", + "caption": "Fig. 2. Generating principle of normal-equidistant trochoids of rack.", + "texts": [ + " 1, the coordinate system Sh is attached to the rack cutter, while xh-axis is collinear with the pitch line of the basic rack. The position vector rh(h) = [xh(h),yh(h)] of each rack segment is listed in Table 1. The unit normal vector of the rack nh(h) = [nxh(h),nyh(h)] can be derived from the following equation: nh\u00f0h\u00de \u00bc k ohrh\u00f0h\u00de jk ohrh\u00f0h\u00dej \u00f06\u00de The normal-equidistant trochoids CD and DE are the specific and significant segments which are deserved to explore. As the generating principle of normal-equidistant trochoids of the proposed rack shown in Fig. 2, the generating process is divided into the following steps: Referring to Figs. 1 and 2, 11 profile parameters, q1, q2, u, v, t, s, j, s, d, c, and ea, are defined on the rack. Parameters q1 and q2, the radii of the circular arcs AB and HI, respectively, are used to define the tooth topland shape of the female rotor; parameters u and v, the pressure angles on the high- and low-pressure sides, respectively, are used to adjust the contact direction near the pitch circle; parameters t and s, the length of the straight lines BC and GH, respectively, are used to modify the length of involutes; and parameters j and s, the normal equidistances of the trochoids, are used to adjust the radii of the sealing arcs on the high-pressure side" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002244_0022-4898(71)90025-5-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002244_0022-4898(71)90025-5-Figure3-1.png", + "caption": "FIG. 3. Coordinate system and notations.", + "texts": [ + " Just before this situation occurs and at the limit when the lateral stress and the smaller principal stress in the vertical plane are equal, failure conditions in the vertical plane govern the stresses; with lateral failure this condition still has to be maintained, otherwise the failure mode would revert to that in the vertical plane. Thus the effect of lateral failure is a limitation on the applied shear stresses; at this limit the stresses can be calculated on the basis of two dimensional failure conditions. The differential equations of plasticity express the combination of the differential equations of equilibrium with the Mohr-Coulomb failure criterion. For the coordinate system and notations shown in Fig. 3, these differential equations are: dz = dx tan (0 4\" ~) da 4, 2a tan q~d0 -- Y [sin (~ 4- q~) dx q- cos (e 4\" q~) dz]. (1) cos q~ The upper sign refers to the family of the slip lines corresponding to the first, and the lower sign corresponds to the second characteristics of the differential equations. For numerical computations, the above differential equations are replaced by four difference equations that allow the computation of the coordinates as well as the ~ and 0 values at a point from values known at two adjacent points previously computed or given at the boundary" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002325_s10846-011-9584-2-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002325_s10846-011-9584-2-Figure2-1.png", + "caption": "Fig. 2 Schematic of C-Plane II vehicle", + "texts": [ + " The aileron-elevon system controls the pitch and roll coupled motion, and the differential speed control of the two rotors also regulates the roll motion whereas the direction of the vehicle is manipulated by the rudder. The vehicle thrust is regulated by the velocity of the propulsion system [12]. 2.2 Modeling The dynamic model for forward flight of this mini VTOL aircraft, considering the aerodynamic effects, is obtained by employing the Euler\u2013 Lagrange formulation. This formulation is introduced as follows. Consider an inertial fixed frame and a body frame fixed attached to the center of gravity of the aircraft denoted by I = {xI, yI, zI} and B = {xB, yB, zB}, respectively (see Fig. 2). The stability frame S = {xS , yS , zS} and the wind frame W = {xW , yW , zW} are considered during the rotation of the wind velocity vector, [22]. Assume the generalized coordinates of the mini UAV as q = (x, y, z, \u03c8, \u03b8, \u03c6)T \u2208 R 6, where \u03be = (x, y, z)T \u2208 R 3 represents the translation coordinates relative to the inertial frame, and \u03b7 = (\u03c8, \u03b8, \u03c6)T \u2208 R 3 describes the vector of three Euler angles with rotations around z, y, x axes. These angles \u03c8 , \u03b8 , and \u03c6 are called yaw, pitch and roll, respectively", + " The norm of vector x is defined as \u2016x\u2016 = \u221a xTx and the induced norm of a matrix is also defined as \u2016A\u2016 = \u221a \u03bbMax{AT A}. The forces acting on the aircraft include those of the propulsion system F p and aerodynamic effects Fa. These forces are described as follows F = F p + Fa with F p = \u239b \u239d Tc 0 0 \u239e \u23a0 , Fa = BB SBB W \u239b \u239d \u2212D Y L \u239e \u23a0 where Tc is the thrust force of the two motors (Tc = T1 + T2). The lift force L, sideforce Y and drag force D are defined as aerodynamic forces, and g is the acceleration due to gravity, (see Fig. 2) [12]. In this analysis, the thrust force is oriented parallel to the axis xB of the body frame. The moments generated on the mini VTOL aircraft are due to actuators (actuator moment act, reaction moment rot and gyroscopic moment gyro), and the aerodynamic effects a. These moments are defined as follows = \u239b \u239d L M N \u239e \u23a0 = act + rot + gyro + a with act = \u239b \u239d \u03c4\u03c6 \u03c4\u03b8 \u03c4\u03c8 \u239e \u23a0 , rot = \u239b \u239d ( Irot1 \u03c9\u0307r1 \u2212 Irot2 \u03c9\u0307r2 ) 0 0 \u239e \u23a0 , gyro = \u239b \u239d 0 r(Ir1\u03c9r1 \u2212 Ir2\u03c9r2) q(\u2212Ir1\u03c9r1 + Ir2\u03c9r2) \u239e \u23a0 , a = \u239b \u239d L\u0304 M\u0304 N\u0304 \u239e \u23a0 where \u03c4\u03c6 = a( fe2 \u2212 fe1), \u03c4\u03b8 = e( fe1 + fe2) and \u03c4\u03c8 = e fr are the control inputs with a and e that represent the distance from the center of mass to the forces fe1 and fe2. \u03c9ri denotes the angular velocity of the rotor, Iri is the inertia moment of the propeller and Iroti is the moment of inertia of the rotor around its axis for i = 1, 2. L\u0304, M\u0304 and N\u0304 are aerodynamic rolling, pitching and yawing moments, respectively (see Fig. 2) [12]. Since the Lagrangian equation contains no crossterms in the kinetic energy combining \u03be\u0307 with \u03b7\u0307, the Euler\u2013Lagrange Eq. 1 can be partitioned into dynamics for \u03be coordinates and \u03b7 coordinates [3]. The translational motion can be obtained using the following expression d dt ( \u2202Lt \u2202 \u03be\u0307 ) \u2212 ( \u2202Lt \u2202\u03be ) = F (3) with Lt = 1 2 \u03be\u0307 T m\u03be\u0307 \u2212 mgz (4) Thus, after some computations, the translation motion of this vehicle is described as \u239b \u239d mx\u0308 my\u0308 mz\u0308 \u239e \u23a0 = RI B \u239b \u239d Tc 0 0 \u239e \u23a0+ RI BBB SBB W \u239b \u239d \u2212D Y L \u239e \u23a0 + \u239b \u239d 0 0 \u2212mg \u239e \u23a0 (5) Similarly, the rotational motion is described as d dt ( \u2202Lr \u2202 \u03b7\u0307 ) \u2212 ( \u2202Lr \u2202\u03b7 ) = (6) where Lr = 1 2 \u03b7\u0307T M(\u03b7)\u03b7\u0307 (7) From Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002322_j.talanta.2012.09.018-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002322_j.talanta.2012.09.018-Figure1-1.png", + "caption": "Fig. 1. (A) Structure of the ionophore. (B) A cell with the holder that used for the absorbance measurements.", + "texts": [ + " Various methods have been proposed for silver detection, including the application of ion-selective electrodes [25,26], flame absorption spectroscopy [27], inductively coupled plasma [28], and stripping voltametry [29]. Furthermore, because of the need to determine low concentration of silver in different samples, the sample preconcentration step has been widely required prior to its analysis [27\u201331]. In the present work, a recently synthesized ionophore [32], 7-(1H-benzimidazol-1-ylmethyl)-5,6,7,8,9,10-hexahydro-2H-1, 13,4,7,10-benzodioxatriaza cyclopentadecine-3,11(4H,12H)-dione, (Fig. 1A) in PVC membrane containing a plasticizer and 3-octadecanoylimino-7-(diethylamino)-1,2-benzophenoxazine (ETH5294) was prepared. A highly selective optical chemical sensor is designed for detection of very low concentrations of silver ions without any pre-concentration steps. The detection limit and selectivity of this optode was comparable to those of all optical sensors mentioned above. 2.1. Chemicals 7-(1H-benzimidazol-1-ylmethyl)-5,6,7,8,9,10-hexahydro-2H-1, 13,4,7,10-benzodioxa-triazacyclopentadecine3,11(4H,12H)-dione, C22H25N5O4 (MW: 423", + "25 was prepared sodium citrate with addition of suitable amounts of nitric acid. A stock solution of 0.010 mol L 1 Ag\u00fe ions was prepared by dissolving an appropriate amount of AgNO3 into a 100 mL standard flask and diluting it to the mark with deionized water. Lower concentrations were prepared by appropriate dilution of the stock solution with citrate buffer of pH 6.25. A double beam UV\u2013vis spectrophotometer (Cary 500 Scan, Varian, Palo Alto, CA, USA) with 1.0 cm quartz cells containing plastic holders (Fig. 1B) was used for the absorbance measurements. pH measurements were taken by the pH/ion meter (Metrohm, Herisau, Switzerland), Model 827, which was equipped with a combined glass electrode. Atomic absorption spectrometer, Shimadzu Model AA670 (Tokyo, Japan), furnished with an Ag-hollow cathode lamp was used. All of the parameters were adjusted according to the standard recommendation of the factory. A typical membrane consisted of 31.0 mg PVC, 62.0 mg DOP, 2.0 mg NaTPB, 2.0 mg ionophore, and 2.0 mg ETH5294 dissolved in 1", + " The mixture was stirred with a magnetic stirrer for 15 min to achieve a homogeneous solution. A glass slide with 9 mm 50 mm was selected and cleaned with 1.0 mol L 1 sulfuric acid and sodium hydroxide solutions, respectively. Then, it was washed with water and dried in an oven at 110 1C for 1 h. The membrane was cast by pipetting a 20 mL aliquot of the membrane solution onto the glass slide, and spreading it rapidly using a capillary glass tube. The membrane was allowed to stand in room temperature to dry for 4 h. The sensor was placed in a 1.0 cm quartz cell (Fig. 1B) containing 3 mL of the citrate buffer solution (pH 6.25). After 150 s, its absorbance was measured at 660 nm. This procedure was repeated for a set of Ag\u00fe standard solutions at different concentrations and the calibration curve was obtained by plotting their absorbance. In this way, the Ag\u00fe concentration contained in the sample can be calculated from the calibration curve. The optode was regenerated in 0.2 mol L 1 HCl solution for 300 s and ready to use. Digestion of Silver Sulfadiazine (1.0% topical cream) was done by burning the cream in an oven and then dissolving its ash in the concentrated HNO3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002050_acc.2010.5531051-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002050_acc.2010.5531051-Figure1-1.png", + "caption": "Fig. 1. 3-DOF Helicopter System.", + "texts": [ + " Robustness to model uncertainties and external disturbances rejection are studied via simulations and experimental results on a 3-DOF Helicopter platform by Quanser. Structure of the paper: The structure of this paper is as follows. Section II contains the system model formulation and the control problem statement. Section III details the proposed observer/controller structures. Section IV and V present some simulations and experimental results, respectively; and Section VI gives some concluding remarks and future works. In this work, a 3-DOF helicopter manufactured by Quanser was employed for the experimental proofs (Fig. 1). The \u201cHelicopter\u201d setup consists of a base on which a long arm is mounted. The arm carries the helicopter body on one end and a counterweight on the other end. The arm can tilt on an elevation axis as well as swivel on a vertical (travel) axis. Quadrature optical encoders mounted on these axes measure the elevation and travel of the arm. The helicopter body, which is mounted at the end of the arm, is free to pitch about the pitch axis. The pitch angle is measured via a third encoder. Two motors with propellers mounted on 978-1-4244-7427-1/10/$26" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001652_j.jsv.2008.12.018-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001652_j.jsv.2008.12.018-Figure8-1.png", + "caption": "Fig. 8. Nodal degrees of freedom for a 3-D beam type finite element.", + "texts": [ + " This lumped beam model has previously been used with good correlation results for previous work ARTICLE IN PRESS Fig. 7. Center bearing stiffness and fit. ~ Measured stiffness and \u2014 fit with: y \u00bc 2341339.0096103x2 0.0000181x+79.6410787, R2 \u00bc 0.9874474. M. Browne, A. Palazzolo / Journal of Sound and Vibration 323 (2009) 334\u2013351340 on similar driveshaft analysis. The lumped mass helps to simplify the FEA model by only having diagonal terms for the mass matrix. Therefore, the FEA model consists of massless beam elements and concentrated inertias as shown in Fig. 8. The nodal rotational and translational degrees of freedom of the 2-noded, 6 degrees of freedom per node beam element in Fig. 8 are arranged in the element displacement vector with the following convention of Ue \u00bc \u00bdxi yi zi yxi yyi yzi xi\u00fe1 yi\u00fe1 zi\u00fe1 yx;i\u00fe1 yy;i\u00fe1 yz;i\u00fe1 T. (4) Subsequently, the diagonal lumped mass matrix and stiffness matrix for the beam element are Me \u00bc diag\u00f0\u00bdmi mi mi Ip;i I t;i I t;i mi\u00fe1 mi\u00fe1 mi\u00fe1 Ip;i\u00fe1 I t;i\u00fe1 I t;i\u00fe1 \u00de (5) ARTICLE IN PRESS M. Browne, A. Palazzolo / Journal of Sound and Vibration 323 (2009) 334\u2013351 341 and Ke \u00bc ae 1 0 0 0 0 0 ae 1 0 0 0 0 0 ae 2 0 0 0 ae 3 0 ae 2 0 0 0 ae 3 ae 4 0 ae 5 0 0 0 ae 4 0 ae 5 0 S ae 6 0 0 0 0 0 ae 6 0 0 Y ae 7 0 0 0 ae 5 0 ae 8 0 M ae 9 0 ae 3 0 0 0 ae 10 M ae 1 0 0 0 0 0 E ae 2 0 0 0 ae 3 T ae 4 0 ae 5 0 R ae 6 0 0 I ae 7 0 C ae 9 2 6666666666666666666666664 3 7777777777777777777777775 , (6) where the coefficients are defined as ae 1 \u00bc EeAe=Le; ae 2 \u00bc 12EeIe x3=L3 e ; ae 3 \u00bc 6EeIe x3=L2 e ; ae 4 \u00bc 12EeIe x2=L3 e ; ae 5 \u00bc 6EeIe x2=L2 e ; ae 6 \u00bc GeJe=Le; ae 7 \u00bc 4EeIe x2=Le; ae 8 \u00bc 2EeIe x2=Le; ae 9 \u00bc 4EeIe x3=Le and ae 10 \u00bc 2EeIe x3=Le: (7215) Initial finite element modeling of the driveshaft with beam elements spaced along the tube, and at every major change of inner or outer diameters produces a system with 162 degrees of freedom" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001355_14644193jmbd187-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001355_14644193jmbd187-Figure3-1.png", + "caption": "Fig. 3 Contact of two bodies with curved surfaces and their principal planes of curvature", + "texts": [ + "857 \u00d7 107 (\u2211 \u03c1o )\u22121/2 (\u03b4\u2217 o) \u22123/2(N/mm3/2) (48b) Then the effective elastic modulus K for the bearing system is written as K = 1 (1/k1/n i + 1/k1/n o )n where n = 3 2 (49) ki and ko are the inner and outer raceways to ball contact stiffness, respectively, and are calculated considering Poisson\u2019s ratio, the dimensionless deflection factor, and the curvature sum at the contact points [33, 34]. When there is an off-sized ball in a bearing, this ball will cause an additional deflection difference. This difference can be larger or smaller than the rest, depending on the off-sized ball diameter. Figure 3 shows that three balls have a greater diameter than the rest of the balls in the set. This ball will force one to squeeze more in relation to the other balls and hence will produce a greater force than the rest. Hence, for this off-sized ball, the equation of displacement of the jth ball becomes (assuming the inner and outer races are rigid) \u03b4\u03b8 = \u03b4\u03b8i + (50) where is the diameter difference of the off-sized ball. Owing to the different ball diameters, the race is deformed into a complex shape that turns with the rotational speed of the cage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000171_iros.2008.4650592-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000171_iros.2008.4650592-Figure1-1.png", + "caption": "Fig. 1: External forces acting on a point mass UGV model.", + "texts": [ + " Since we seek the time optimal trajectory, the global search selects paths along which the UGV can sustain high speeds without violating dynamic constraints such as rollover, excessive side slip, and maintaining ground contact. Velocity limits (above which some of the dynamic constraints may be violated) are computed by mapping the dynamic constraints to constraints on the vehicle's speed and tangential acceleration. For long range planning, the UGV is modeled as a suspended point mass [2, 18]. For short range planning, the vehicle is modeled as a rigid body [9]. Velocity limits for a point mass model are derived by first expressing the three external forces shown in Fig. 1 in terms of the UGV\u2019s speed and tangential acceleration (see [2,18] for a detailed derivation): 2 2 snmmgkR snmmgkf smmgkf rr qqq tt & & && \u03ba \u03ba += += += (1) where ft and fq are components of the friction force tangent and normal to the path, \u03ba is the path curvature, k is a unit vector pointing opposite of the gravity force, n is a unit vector pointing in the direction of the path center of curvature, and the subscripts denote projections along the path coordinate frame, t, q, r. We can now express the dynamic constraints in terms of the external forces (1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001193_j.triboint.2008.04.002-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001193_j.triboint.2008.04.002-Figure5-1.png", + "caption": "Fig. 5. Contact geometry and surface contact area.", + "texts": [ + " It was noticed that the grease structure did not suffer too much with the stress cycles to which has been submitted, and could be considered that it has not aged. The discs were manufactured in AISI 52100 steel, a very common material used for all types of bearings. The discs dimensions are 70 mm of diameter and 7 mm wide. In order to avoid a stress raise due to edge effects, as a result of an eventual misalignment, one of the discs was kept cylindrical (cylindrical disc) while the other was manufactured with a transversal radius of 35 mm (spherical disc). Consequently, when loaded against each other, an elliptical area is generated (see Fig. 5). The discs were loaded to achieve a peak contact pressure of 2 GPa. Due to the geometry of the discs and according to the Hertzian contact theory (see for instance [21]), this results in a contact ellipse with the dimensions: a \u00bc 0.52 mm and b \u00bc 0.83 mm. The discs rolled under pure rolling conditions at a speed of 3000 rpm. For all the tests, using oil or grease, a permanent feeding of the lubricant in the inlet region exists, ensuring fully flooded conditions. Various authors [22\u201324] tested different types of artificial defects such as conical, spherical and diamond dents, among e; (b) apparatus for grease lubricated contacts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001491_09544062jmes1452-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001491_09544062jmes1452-Figure12-1.png", + "caption": "Fig. 12 Area efficiency of other common designs", + "texts": [ + " On other lectures of claw rotor design, the rotor profile is designed by the theory of gearing that including the coordinate transformation and meshing equation. The method can yield the conjugate profile with noninterference. However, it should be derived from more complicated rotors equations and to solve the meshing equation in order to obtain the completed rotor profile. The following shows two cases in which the rotor profile is produced by the method above. The area efficiency is calculated as 44.78 per cent (Case 1) and 46.45 per cent (Case 2) by considering the carryover area (as shown in Fig. 12). In this article, the area efficiency with considering carryover area is shown in Fig. 11. The results show that the area efficiency is near the common designs. However, this article presents a simple mathematical model without the coordinate transformation and meshing equation. The article does not need coordinate system for coordinate transformation operation. The design method in this article is much simpler than other designs. On the other hand, when assembling the two rotors, a tiny clearance should be kept between the two rotors and from the rotor to the chamber for avoiding noise and heating problem caused by direct contact friction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001789_1.4002165-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001789_1.4002165-Figure8-1.png", + "caption": "Fig. 8 Generating surfaces for \u201ea\u2026 convex and \u201eb\u2026 concave sides", + "texts": [ + " Condition of parallelism of vectors P2 and OP is given by \u2212 2 cos R2 Am cos 2 = P \u2212 2 sin R2 \u2212 Am sin 2 7 hich yields P 2 = sin 2 8 cos 2 ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 03/17/201 Such relation is satisfied since point P belongs to the instantaneous axis of rotation by cutting of surfaces P and 2. Since the instantaneous axis of rotation by cutting of surfaces P and 2 coincides with the instantaneous axis of rotation by meshing of surfaces 1 and 2, this means that surfaces P, 2, and 1 will be in tangency at reference point P. Point P would coincide with point M defined in Fig. 4 b , 5 a , and 5 b . Figure 8 shows the generating surfaces corresponding to the convex and concave sides. The generating surface is a cone obtained by rotation of the edge of the blade. This rotation is not related with process of generation but with the required cutting velocity. The surface parameters are u , , wherein u is the profile parameter and is the longitudinal parameter. The generating surface is defined basically with two magnitudes see Fig. 8 , the blade angle gi i=1,2 and the cutter point radius Rgi i=1,2 . Subscript i=1 is applied to the convex side and subscript i=2 is applied to the concave side. Cutter point radii are given by Rg1 =OgP1 and Rg2=OgP2 see Fig. 7 . Since the generating surface is a revolution surface, the cutter radius depends uniquely on profile parameter u as rc u = Rgi u sin gi, i = 1,2 9 wherein the upper sign is applied to the convex side and the lower sign is applied to the concave side. This criterion for the sign is considered in the following derivations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure7-1.png", + "caption": "Fig. 7 Configuration 4 of the example", + "texts": [ + " 5 \u2013 7 , n=n0 \u2212n\u22121=5\u22121=4, p=n1=4, and F=3 n\u22121 \u22122p=3 4\u22121 \u22122 4 =1. 3 If Link 3 is attached together with Link 4 by a pin, P, as shown in Fig. 6 a , the revolute pair between Link 3 and Link 4 is then frozen. The mechanism is termed as Configuration 3 and is shown in Fig. 6 b . If this configuration comes from Configuration 2, the adjacency matrix A3 can be obtained by multiplying \u22121 on A2 2,3 , A2 3,2 , A2 3,4 , and A2 4,3 , i.e., A3 = 0 1 0 0 1 1 0 1 0 0 0 1 0 \u2212 1 0 0 0 \u2212 1 0 1 1 0 0 1 0 10 4 If Link 4 is attached to Link 5 by a pin, P, as shown in Fig. 7 a , the revolute pair between Links 4 and 5 is then frozen. The mechanism is termed as Configuration 4 and is shown in Fig. 7 b . If this configuration comes from Configuration 3, the adjacency matrix A4 can be obtained by multiplying \u22121 on A3 3,4 , A3 4,3 , A3 4,5 , and A3 5,4 , i.e., A4 = 0 1 0 0 1 1 0 1 0 0 0 1 0 1 0 0 0 1 0 \u2212 1 1 0 0 \u2212 1 0 11 ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash 5 If Link 1 is attached together with Link 5 by a pin, P, as shown in Fig. 8 a , the revolute pair between Links 1 and 5 is frozen. The mechanism is termed as Configuration 5 and is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002277_icit.2012.6210043-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002277_icit.2012.6210043-Figure4-1.png", + "caption": "Fig. 4. Flat-Earth Inertial and Body Fixed Frame", + "texts": [ + " The tilting of the tail rotor is not utilized in the yawing motion control as in typical Tri-Rotor designs, but in the decoupling of the rolling motion from the pitching tail thrust. This is achieved by controlling the tail rotor's rotation angle via the tail servo so as to track the rolling angle and produce an opposite and equal rotation angle. Through this, we ensure that the tail rotor thrust is always applied as a pitching motion component. The aforementioned concepts are presented in Figure 3. Let B = {Bx, By, Bz} be the coordination body-fixed reference frame and E = {Ex, Ey, Ez} be the Flat-Earth model inertial reference frame as depicted in Figure 4. It should be noted that the earth inertial frame (EFF) follows the North East-Down (NED) notation and the body-fixed inertial frame (BFF) follows the standard aircraft notation where the z axis points downwards, the x-axis towards the longitudinal flight direction and the y-axis towards the right wing. Also let U = {u, v, w} be the vector of linear rates and Q = {p, q, r} the vector of angular rates expressed on the coordination frame B and let X = {x, y, z} be the vector of translational displacements and e = {/{), e, lJI} the vector of rotational displacements expressed on the inertial frame E" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000060_1.2918917-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000060_1.2918917-Figure3-1.png", + "caption": "Fig. 3 Illustration of a planar linkage mechanism with five links", + "texts": [ + " Suppose Fm=2, as the example given bove, then changing any pair of elements of the original adjaency matrix symmetrical from value 1 into value \u22121 will give a ifferent configuration of the metamorphic mechanism. If Fm=3, hanging any two pairs of elements of the original adjacency marix from value 1 into value \u22121 will also give a different configuation of the metamorphic mechanism, and so on. ig. 1 Illustration of five-bar spherical metamorphic echanism ig. 2 The configuration when a five-bar linkage configuration hanges to a four-bar linkage configuration 74501-2 / Vol. 130, JULY 2008 om: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash 3 Examples of Planar Mechanisms Figure 3 shows a planar mechanism with five links. The original adjacency matrix of the mechanism, A0, is the one in Eq. 1 with Fm=2. There are five more possible configurations of the mechanism as there are five 1 among upper triangle of A0. Each one will result in a different adjacency matrix as shown below. 1 If Slider 2 is fixed to Link 1, as shown in Fig. 4 a , the prismatic pair between Link 1 and Slider 2 is frozen. The mechanism is termed as Configuration 1 and is shown in Fig. 4 b . It works as a general four-bar linkage mechanism with DOF F=1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000248_s0022-0728(79)80335-6-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000248_s0022-0728(79)80335-6-Figure2-1.png", + "caption": "Fig. 2. Dependence of the exchange current I 0 of EAR on electrode potential e (dimensionless quantities) for g = 0 and various values of transfer coefficient/3.", + "texts": [ + " Explicit formula may be obtained only when e is eliminated: This dependence does not contain the heterogeneity factor g, therefore it is common for Langmuir and Frumkin isotherms. For the Langmuir isotherm eqn. (29) may be transformed into an explicit dependence of exchange current on potential: which in the particular case of fl = 0.5 transforms into: Io = I o ( e ) curve: only for fi = 0.5 the curve is symmetrical with respect to the potential of its maximum. For f i \u00a2 0.5 the curve is asymmetrical, with its maximum shifted towards negative values if fi < 0.5 and towards positive values if fl > 0.5. This point is illustrated by Fig. 2, showing the Io = Io (e ) dependences for Langmuir isotherm and various values of transfer coefficient. The maximum value of I0 for every value o f g is given by: Io,max = (1 __fl)l--~. f~ (32) The coverage corresponding to this maximum exchange current density: 0max = 1 --fl (33) and the corresponding potential: emax = In {fi/(1 --/3) } -- g(1 --/3) (34) The dependences Io = lo(e) forg ~ 0 may be computed numerically from (30) 11 Fig. 3. Dependence of the exchange current I 0 of EAR on electrode potential e (dimensionless quantities) for ~ = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001792_s11249-010-9617-1-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001792_s11249-010-9617-1-Figure2-1.png", + "caption": "Fig. 2 a Two asperities with the same radius of curvature; b asperities in contact", + "texts": [ + " Thereby, the coefficient of friction is infinitely large, because there is already a finite frictional force in the absence of a normal force. Here, we want to investigate the non-adhering surface, meaning that the condition dc \\ l is met.1 Furthermore, we assume that the typically met conditions, dc/R 1 and l/R 1, are valid, which justifies the use of half-space theory. First, we investigate a contact between two asperities and, thereafter, perform an averaging over the statistical distribution. In Fig. 2, one asperity from each surface is shown and parameterized. The form of the asperities is described by z1\u00f0x\u00de \u00bc Z1 x2 2R ; z2\u00f0x\u00de \u00bc Z2 \u00fe x X0\u00f0 \u00de2 2R : \u00f02\u00de As long as the surfaces remain in contact, the penetration depth is given by d \u00bc Z1 Z2 X2 0 4R ; \u00f03\u00de and the tangent of the contact angle h (see Fig. 2b) by tan h h X0= 2R\u00f0 \u00de: \u00f04\u00de The first and last contact take place when the conditions d = 0 or d = -dc are met, respectively. From this, we obtain X0;min \u00bc 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R Z1 Z2\u00f0 \u00de p ; X0;max \u00bc 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi R Z1 Z2\u00f0 \u00de \u00fe Rdc p : \u00f05\u00de If we denote the force of interaction between the asperities as F, then the z-component of the force FN and the xcomponent of the force FR, neglecting higher order terms of h, are given by FN F; FR FX0= 2R\u00f0 \u00de: \u00f06\u00de The coefficient of friction can then be calculated as l \u00bc FR = FN ; where the notation " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003565_1350650112468071-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003565_1350650112468071-Figure1-1.png", + "caption": "Figure 1. Structure of first generation bump foil bearing.", + "texts": [ + " Thick plate model and Kirchhoff plate model of top foil are established, the finite element method and finite difference method are coupled together to solve Reynolds equation and foil structural deflections. Minimum film thicknesses along bearing length direction and film thickness in circumferential direction at bearing mid-plan of these two models are compared. The results indicate that the shear stiffness prevents axials and circumferential variations in gas film thickness. Menawhile, the shear effects are necessary in modeling 2D foil structual models for journal bearing. Bump-type foil bearing structure description The structure of the first generation bump foil bearing is shown in Figure 1. It consists of a thin top foil bump and a series of corrugated bump strip layer which depicted in Figure 2. The leading edges of both bump and top foil are spot welded to the bearing sleeve, the trailing edges of foils are free. The bump foils act like springs. As the shaft rotates, the top foil would deflect when air pressure forces act on its smooth surface. In 2D coordinates, the continuity equation and Navier\u2013Stokes equations can be written as @u @x \u00fe @w @z \u00bc 0 \u00f01\u00de u @u @x \u00fe w @u @z \u00bc @p @x \u00fe @2u @x2 \u00fe @2u @z2 \u00f02\u00de u @w @x \u00fe w @w @z \u00bc @p @z \u00fe @2w @x2 \u00fe @2w @z2 \u00f03\u00de To make the analysisi of fluid film bearings with thin-film situation relatively simple, Reynolds based his theory of lubrication and derived Reynolds equation from Navier\u2013Stokes equation by using some assumptions (Reynolds, 1886)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000559_j.triboint.2007.07.009-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000559_j.triboint.2007.07.009-Figure5-1.png", + "caption": "Fig. 5. Detail drawing of new test specimen (dimensions in mm).", + "texts": [ + " The center of the sphere used in the current study was contained within the upper specimen by roughly 1mm. This dimension is not critical as long as the contact surface on each specimen is perpendicular to its centerline within reasonable manufacturing limits of approximately 0.11. The clearance between the alignment pin and the specimens was approximately 0.03mm. The spherical portion of the alignment pin is commercially available from industrial supply companies and is called a \u2018\u2018tooling ball\u2019\u2019. A drawing depicting the critical and major features of the specimens is shown in Fig. 5. There are several features of the specimens that are worth noting at this time. The specimens were produced using 15.9mm (5/8 in.) hex stock. The hex provided a convenient means of gripping the specimens during manufacturing and testing. The cylindrical section and the contact face were machined using a CNC lathe. The toolpath of the lathe was programmed such that the 12.70mm (1/2 in.) diameter and the contact face were cut with the tool in continuous contact with the specimen. This eliminated any burrs on the corner of the 12" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001095_s11044-010-9236-5-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001095_s11044-010-9236-5-Figure2-1.png", + "caption": "Fig. 2 Beam element of the absolute nodal coordinate formulation", + "texts": [ + " In the finite element absolute nodal coordinate formulation, the global position vector ri of an arbitrary point on element i can be defined as [4] ri = Si (x\u0304i )ei (1) where Si (x\u0304i ) is the element shape function matrix, x\u0304i = [xi yi zi]T is the vector of spatial coordinates defined in the element coordinate system, and ei is a vector of nodal coordinates of element i. The nodal coordinate vector eik at node k of element i is defined as follows: eik = [ (rik)T ( \u2202rik \u2202xi )T ( \u2202rik \u2202yi )T ( \u2202rik \u2202zi )T ]T (2) where rik is the global position vector at node k; \u2202rik/\u2202xi , \u2202rik/\u2202yi and \u2202rik/\u2202zi define the gradient vectors of the global position vector at node k along the element on x, y and z axes as shown in Fig. 2. Elements parameterized using all the 3 sets of gradient vectors are called fully parameterized elements of the absolute nodal coordinate formulation. Equation (1) can be expressed for a beam element as ri = ri c + yi \u2202ri c \u2202yi + zi \u2202ri c \u2202zi (3) where ri c is the global position vector on the beam centerline and \u2202ri c/\u2202y i and \u2202ri c/\u2202zi are the transverse gradient vectors that can be used to describe the orientation and deformation of cross section. These vectors can be, respectively, interpolated as [7] ri c = Si I ( xi ) ei , \u2202ri c \u2202yi = Si II ( xi ) ei , \u2202ri c \u2202zi = Si III ( xi ) ei (4) where Si = Si I +yiSi II +ziSi III" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002365_elan.201000427-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002365_elan.201000427-Figure3-1.png", + "caption": "Fig. 3. Effect of the current intensity and deposition time of gold on the analytical signal. B-AP concentration, 1.0 10 10 M. 3-IP concentration, 1.0 mM; silver ion concentration, 0.4 mM; BSA 1%. SPCE-MWCNTs (0.1 mgmL 1 MWCNTs) and AuCl4 (1.0 mM).", + "texts": [ + " Gold nanoparticles have been formed from AuCl4 of different concentrations prepared in 0.1 M HCl by applying a constant current intensity for a period of time. Three gold concentrations (0.1 mM AuCl4 , 0.5 mM AuCl4 , 1 mM AuCl4 ), 3 current intensities ( 5 mA, 10 mA, 100 mA) and 3 deposition times (60 s, 120 s and 300 s) were studied using the procedure described in Section 2.3.2. 66 www.electroanalysis.wiley-vch.de 2011 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim Electroanalysis 2011, 23, No. 1, 63 \u2013 71 Figure 3 shows the results obtained for the gold concentration of 1 mM with the transducer previously modified with MWCNTs. In all cases it can be observed that the analytical signal obtained is similar for all times assayed. Regarding the three current intensities applied ( 5, 10, 100 mA) there is no considerably difference from 5 to 10 mA since the slightly enhancement of the current can be justified with the standard deviation. As was reported in a previous work from Mart nez-Paredes and colleagues [19] the formation of gold nanoparticles is influenced by the current intensity that is applied, the time of gold deposition and gold concentration. Namely, the particle diameter seems to be indirectly proportional with time deposition and reduction intensity. For large deposition times occurs a shift in potential towards more negative potentials ( 0.70 V) during gold electrodeposition. In the acid medium the generation of hydrogen occurred at this potential improving the nucleation of gold on the electrode surface in a detriment of growth of nanoparticles. The bar diagrams in Figure 3 shows that current decreases when increases the time for electrodeposition. For a 60 s deposition time the best analytical signal was achieved applying a constant current intensity of 100 mA. However, the results revealed that for these experimental conditions ( 100 mA) there was a considerable standard deviation. Instead, a current intensity value of 5 mA was chosen for further studies. These results were also observed for the other gold concentrations in study (data not show). In order to achieve the best relation between sensitivity and reproducibility the concentration of 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002804_tro.2011.2181098-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002804_tro.2011.2181098-Figure7-1.png", + "caption": "Fig. 7. Deformable object model. (a) Object model. (b) Three-DOF joint unit. (c) Contact model.", + "texts": [ + " 2) The plate\u2019s surface area is larger than that of the object. 3) The object is deformable, and its thickness is small. 4) The object is isotropic, and it has uniform mass distribu- tion and uniform viscoelasticity. 5) The nominal pressure distribution on the object is uniform. 6) The friction coefficient between the plate and the object that is based on Coulomb\u2019s law is uniform and is given by \u03bcs and \u03bck for static and dynamic coefficients, respectively. Deformable object model: For a thin deformable object, we consider virtual tile links as shown in Fig. 7(a). The link is a square with sides of length l. Based on the shape and the size of the modeled object, the arrangement of virtual tiles is determined. A node with a mass of m is located at the center of the link, where neighboring nodes are connected to each other by a viscoelastic joint unit as shown in Fig. 7(b). The joint unit is composed of three DOFs: bending, compression/tension, and torsion. The bending and the compression joints have viscoelastic elements that are given by a Kelvin\u2013Voigt model, while the torsion joint is free for simplicity of the simulation model. In Fig. 7(b), kb and cb express the elasticity and viscosity, respectively, of the bending joint. Similarly, kc and cc are the elasticity and viscosity, respectively, of the compression joint. Contact model: Fig. 7(c) shows the contact model between the plate and the ith virtual link. The contact force is computed with the penalty method that is based on the Kelvin\u2013Voigt model [27]. The contact force f contact i that is applied to the node is given by f contact i = kcontacta 2.2 i + ccontact a\u0307i (ai \u2265 0) (1) where ai , kcontact , and ccontact are the distance between the surface of the plate and that of the virtual link, the elasticity, and the viscosity, respectively. In addition, the frictional force f friction i that is applied to the node is given by f friction i = \u03bc\u2217f contact i ", + " Since, in our model, we want to describe the viscoelasticity over the lateral surface of the link (ld), we convert p\u0302s to obtain p\u0302c = [c\u0302c k\u0302c ]T as p\u0302c = p\u0302s (d/l)2 (9) which expresses the scaled viscoelasticity for a contact force that is applied to the lateral surface of the link with area ld and thickness l, based on assumption 4 that the object is isotropic. As a real deformable object, a slice of cheese is employed in the experiment since it is an artificial product that can reasonably correspond to assumptions 4, 5, and 6. Based on the model shown in Fig. 7(a), each squared link has a length l = 10 mm, thickness d = 2.5 mm, and mass m = 0.285 g. To estimate the viscoelastic parameters in bending, a slice of cheese was cut to get the two-link model shown in Fig. 8(a). Actually, its total length is 30 mm, and its width is 10 mm, as shown in Fig. 9(a), where the left 10 mm of the object was gripped at the wall portion. We placed three red markers: the first one at the wall boundary, the second one at the middle of the right 20 mm of the object, where the virtual joint is located, and the third one at the right tip of the object, which is left free" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003379_s13738-011-0002-2-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003379_s13738-011-0002-2-Figure2-1.png", + "caption": "Fig. 2 Cyclic voltammograms of (a) CPE in 0.1 M phosphate buffer solution (pH 8.0) at scan rate 10 mV s-1 and (b) as (a) ?0.2 mM LD; (c) as (a) and (d) as (b) at the surface of BNH/TiO2/CPE and TiO2/ CPE, respectively. Also, (e) and (f) as (b) at the surface of BNH/CPE and BNH/TiO2/CPE, respectively, and (g) as (a) at the surface of BNH/CPE", + "texts": [ + " Electrochemical properties of BNH/TiO2/CPE An investigation of the electrochemical properties and, in particular, the electrocatalytic activity of BNH in aqueous Scheme 1 Structure of levodopa (a) and carbidopa (b) media was conducted by cyclic voltammetry. This compound is insoluble in aqueous media; therefore, we prepared BNH/TiO2/CPE and studied its electrochemical properties in a buffered aqueous solution (pH 8.0) using cyclic voltammetry. The cyclic voltammograms of BNH on the BNH/TiO2/CPE and BNH/CPE exhibited an anodic and the corresponding cathodic peak (Fig. 1, and curve g of Fig. 2, respectively). Experimental results show welldefined and reproducible anodic and cathodic peaks (with Epa = 0.18 V, Epc = 0.09 V, E1/2 = 0.135 V versus SCE and DEp = 0.09 V) for BNH; therefore, this substance can be used as a mediator for the electrocatalysis of some important biological compounds with slow electron transfer. The effect of TiO2 nanoparticles on the electrochemical behavior of BNH can be seen from curve c and curve g in Fig. 2. The peak current of the BNH molecule increased, but the potential did not change in the presence of TiO2 nanoparticles. The peak separation potential, DEp = (Epa\u2013Epc), was greater than the 59/n mV expected for a reversible system. This suggests that the redox couple in BNH/TiO2/CPE shows quasi-reversible behavior in an aqueous medium. In addition, the effect of the potential scan rate on the electrochemical properties of the BNH/TiO2/CPE was studied in an aqueous solution with cyclic voltammetry (Fig", + " A potential-pH diagram was constructed by plotting the calculated E1/2 values as a function of pH. This diagram is composed of a straight line with slope = 50.61 mV/pH. Such a behavior suggests that it obeys the Nernst equation for a two-electron and protontransfer reaction [32]. Electrocatalytic oxidation of LD at a BNH/TiO2/CPE The cyclic voltammetric responses from the electrochemical oxidation of 0.2 mM LD at the BNH/TiO2/CPE (curve f), BNH modified CPE (BNH/CPE) (curve e), TiO2 nanoparticles CPE (TiO2/CPE) (curve d) and unmodified CPE (curve b) are shown in Fig. 2. As shown, the anodic peak potential for LD oxidation at the BNH/TiO2/CPE (curve f) and BNH/CPE (curve e) was about 180 mV, while at the TiO2/CPE (curve d) the peak potential was about 550 mV. At the unmodified CPE, the peak potential was about 590 mV of LD (curve b). Based on these facts and figures, it was concluded that the best electrocatalytic effect for LD oxidation was observed at the BNH/TiO2/CPE (curve f). For example, results show that the peak potential of LD oxidation at the BNH/TiO2/CPE (curve f) shifted by about 370 and 410 mV toward negative values when compared with those of the TiO2/CPE (curve d) and unmodified carbon paste electrode (curve b), respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003599_cca.2011.6044516-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003599_cca.2011.6044516-Figure1-1.png", + "caption": "Fig. 1. Airplane frontal view.", + "texts": [ + " Notation: Symbol R is used to denote the sets of real numbers and sign(z), with z \u2208 R, is the signum function defined as sign(z) = 1 if z > 0 0 if z = 0 \u22121 if z < 0. The model that describe the PVTOL including the actua- tors dynamics is given by [5], [12] x\u0308 = \u2212 sin(\u03b8)z1 + \u03b5z2 + wx (1) y\u0308 = cos(\u03b8)z1 + \u03b5z2 sin(\u03b8) \u2212 1 + wy (2) \u03b8\u0308 = z2 + w\u03b8 (3) \u00b51z\u03071 = \u2212z1 + u1 (4) \u00b52z\u03072 = \u2212z2 + u2. (5) In the above equations, x(t) \u2208 R is the horizontal position, y(t) \u2208 R is the vertical position, \u03b8(t) \u2208 R is the angle with respect to the horizontal reference, z1(t) \u2208 R is the total force, z2(t) \u2208 R is the outward applied force (see Fig. 1), w = (wx, wy, w\u03b8) are the external disturbances affecting the system, and t \u2208 R is the time. Equations (4)\u2013(5) represent the dynamics of the actuators with inputs u1(t) and u2(t) where \u00b51 and \u00b52 are positive constants that define the time 978-1-4577-1063-6/11/$26.00 \u00a92011 IEEE 1482 response of z1(t) and z2(t). The constant \u03b5 is a very small coefficient and difficult to know related to the rotation and lateral acceleration of the airplane such that can be defined as zero [13]. The constant -1 is the normalized gravitational acceleration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003460_1350650112466768-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003460_1350650112466768-Figure2-1.png", + "caption": "Figure 2. Forces and moments acting on the piston.", + "texts": [ + " So, a comprehensive lubrication model should be built to evaluate the influence of various deformations of the piston skirt and the liner on the piston\u2013liner system lubrication. As an effort to reach this, it will be discussed in details that the calculation of the deformations and the influences of the deformations on the piston secondary motion, the oil film pressure and thickness, and the friction loss of the piston skirt\u2013 liner system in this study. A schematic drawing of the piston\u2013liner system is given in Figure 1, and all the forces and moments acting on the piston are shown in Figure 2. The equations of motion are derived based on the dynamic equilibrium of all the forces and moments applied to the piston. The system equation can be expressed as the following mpin\u00f01 a L\u00de \u00fempis\u00f01 b L\u00de mpin a L\u00fempis b L Ipin pis L \u00fempis\u00f0a b\u00de\u00f01 b L\u00de mpis\u00f0a b\u00de bL I pin pis L 2 4 3 5 \u20acet \u20aceb \" # \u00bc F \u00fe Fs M\u00feMs \" # \u00f01\u00de where Fs \u00bc X3 i\u00bc1 Ffi \u00f0Fg \u00fe Fip \u00fe Fic \u00fe Fr\u00feFf \u00de sin \u00fe fp sgn\u00f0 _ \u00de cos cos fp sgn\u00f0 _ \u00de sin \u00f02\u00de Ms \u00bcMf\u00feFgCp FicCg\u00fe X3 i\u00bc1 FfiLi \u00feFrCp \u00f0Fg\u00fe Fic\u00fe Fip\u00feFr\u00feFf \u00de fp sgn\u00f0 _ \u00de rpin cos fp sgn\u00f0 _ \u00de sin \u00f03\u00de fp is the friction coefficient of the wrist-pin bearing", + "comDownloaded from Cg horizontal distance between piston center of mass and piston pin Co crankshaft offset d( , y, t) increase of gap between two surfaces due to deformations of skirt and liner eb, et distance between the skirt center and the cylinder axis at the bottom and top of the skirt, respectively fskt ( , y) piston skirt profile F total side force of piston skirt due to lubrication and contact FB connecting rod force acting on piston Ff total friction force of piston skirt due to lubrication and contact Ffp friction force at the wrist-pin bearing Ff1 friction force at the first piston-ring groove Ff2 friction force at the second pistonring groove Ff3 friction force at the third piston-ring groove FG combustion gas force acting on the top of piston Fic, Fip lateral inertia force of piston and piston pin due to secondary motion of piston Fic, Fip reciprocating inertia force of piston and piston pin Fr total normal force acting on the piston-ring groove Fs force, intermediate variable h oil film thickness h\u0302 convection coefficient Ipis rotary inertia of piston about its center of mass I pin pis rotary inertia of piston about the piston pin, I pin pis \u00bc Ipis \u00fempis\u00bd\u00f0a b\u00de2 \u00fe C2 g k heat conduction coefficient lc length of connecting rod L length of piston skirt mpin mass of piston pin mpis mass of piston M moment of F about the piston pin Mf moment of Ff about the piston pin Mpis inertial moment of piston Ms moment, intermediate variable p oil film pressure pc contact pressure r crankshaft radius R piston radius T0 initial temperature Ttdc,Tmid,Tbdc liner temperature at TDC, middle stroke and BDC of the piston skirt Tf temperature of the surroundings u sliding velocity of piston X, Y global coordinate system as shown in Figure 2 connecting rod angle tilt angle of piston , G Lame\u0301 coefficients lubricant viscosity , y local coordinate system on the piston skirt , T1, T2 parameters in the Vogel equation f friction coefficient of asperity contact fp friction coefficient of the wrist-pin bearing 1, 2 Poisson\u2019s ratios of piston and liner density of oil standard deviation of combined roughness of skirt and liner shear stress \u2019 crank angle _\u2019 angular speed of crankshaft c contact factor fs, fp shear stress factors s shear flow factor x, y pressure flow factors at UNSW Library on July 23, 2015pij" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001374_s12206-009-0101-5-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001374_s12206-009-0101-5-Figure3-1.png", + "caption": "Fig. 3. Mesh generation for heat flux and thermal analysis of induction heating.", + "texts": [ + " To handle this constant change, a quasi-stationary state to the traveling direction of the inductor is assumed for the heat flux and heat flow analyses. Then, the heat-flux generation and heat flow can be modeled with a moving coordinate system, which has the origin at the position of the inductor. The governing equation of heat flow in the moving coordinates is formulated for the FEM and the formulated equation is then implemented to a FORTRAN program to obtain a temperature distribution in the steel plate during the induction heating process [8]. The mesh division for the analysis is shown in Fig. 3, which consists of 3-dimensional 8-node elements. In the developed program, heating speed, current and frequency are set as the input parameters. Fig. 4 illustrates the procedure for analyzing the electro-magnetic field and heat flow. In the analyses, iterations are carried out at each step with updated material properties until the convergence limit by comparing the temperature distributions of subsequent iteration steps. The comparisons finally produce a heat flux at each node from the inductor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002715_j.triboint.2013.03.005-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002715_j.triboint.2013.03.005-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of radial lip seal.", + "texts": [ + " Therefore, the aim of this study is to investigate the effects of macroscopic structure, material parameters and assembly conditions Nomenclature Di shaft diameter E Young's modulus of elastomeric ring F F-test result Fs spring force f friction coefficient h nominal film thickness, separation between surface means h\u0302 dimensionless film thickness, h=s h\u0302T dimensionless local film thickness, truncated if in contact, hT=s ^hT average local film thickness, equals h\u0302 if not in contact hs static undeformed film thickness In influence coefficient for normal deformation Is influence coefficient for circumferential deformation K aspect ratio of solution space, Lx=Ly Lx length of solution domain in x direction Ly length of solution domain in y direction (contact width) pa ambient pressure pc contact pressure pcav cavitation pressure pref characteristic reference pressure ps sealed pressure psc static contact pressure p\u0302f dimensionless fluid pressure, pf =pref q\u0302y dimensionless pumping rate in y direction per unit length in the x direction Q\u0302y dimensionless pumping rate in y direction Q reverse pumping rate R asperity radius r lip dimension, Fig. 1 S cavitation index s lip dimension, Fig. 1 t lip dimension, Fig. 1 T frictional torque x\u0302 dimensionless circumferential coordinate, x=Lx y\u0302 dimensionless axial coordinate, y=Ly Greek letters \u03b1 oil-side seal angle \u03b2 air-side seal angle \u03b4 interference \u03b4n normal deformation of lip surface \u03b4s circumferential deformation of lip surface \u03b7 asperity density \u03bbx autocorrelation length in x direction \u03bby autocorrelation length in y direction \u03be dimensionless number, \u03bcULx=pref s2 \u03c1f density of the full film (uncavitated) \u03c1\u0302 dimensionless lubricant density, \u03c1=\u03c1f s the standard deviation of surface heights \u03c4\u0302avg dimensionless average viscous shear stress in the x direction, \u03c4avg=E \u03c4f shear stress due to contacting asperities \u03d5c:c dimensionless density flow factor \u03d5f , \u03d5f ss, \u03d5f pp dimensionless shear stress factors \u03d5s:c:x, \u03d5s:c:y dimensionless shear flow factors \u03d5xx, \u03d5xy dimensionless pressure flow factors \u03d5yx, \u03d5yy dimensionless pressure flow factors \u03a6 variable representing pressure/average density, defined by Eq", + " The L50\u00f0511\u00de orthogonal array (described in Section 4) is used to construct a numerical experiment plan with eight parameters. The significance of each parameter on the reverse pumping rate and friction torque of the radial lip seal is evaluated, and optimal values of the parameters are suggested. The simulated characteristics of the seal with the set of optimal values are compared with those of the previously-designed seal, mentioned above. 2. Structure and parameters of radial lip seal A typical example of a radial lip seal is shown in Fig. 1. It consists of an elastomeric ring, steel frame and a garter spring. The seal is mounted on a polished shaft with an interference fit. Due to the difference between the sealing lip inner diameter and the shaft diameter (interference), \u03b4, and spring force Fs, a static contact pressure distribution is generated between the seal and the shaft. When the shaft rotates, the microstructures on the lip surface are skewed so as to pump oil toward the center of the contact zone. Because the oil-side angle of the sealing lip \u03b1 is designed to be greater than the air-side angle \u03b2, the tangential deformation of the lip surface will be non-symmetrical so that a net pumping flow is obtained from the air-side to the oil-side of the seal to prevent the leakage", + " The reverse pumping rate Q and friction torque T are often used to evaluate whether the seal is successfully designed. Research and development over many years has revealed that the reverse pumping rate and frictional torque depend on the macroscopic structure and detailed sizes, elastomeric material properties and assembly condition of the seal, besides the microstructure of the seal surface (i.e. asperity characteristics). In this study, the influences of eight parameters on the reverse pumping rate are investigated. The definitions of the eight parameters are shown in Table 1 and Fig. 1. Numerical approaches have often been employed to study the sealing behavior of the radial lip seal. A mixed elastohydrodynamic lubrication model that was previously constructed by the present authors is used in the present study [17]. The model consists of a coupled hydrodynamic lubrication analysis, asperity contact analysis, and deformation analysis, with an iterative computational procedure. It is outlined below. The fluid mechanics of the lubricant film in the sealing zone is governed by the Reynolds equation. Noting the film thickness is much smaller than the seal radius, a Cartesian coordinate system is used. The x direction represents the circumferential direction, and the y direction represents the axial direction, as shown in Fig. 1. Flow factors are used to take into account the effect of the lip surface roughness, while the shaft is assumed to be perfectly smooth. As cavitation is expected, a form of the Reynolds equation that accounts for cavitation is used, as shown in Eq. (1). \u2202 \u2202x\u0302 h\u0302 3 \u03d5xx \u2202\u00f0S\u03a6\u00de \u2202x\u0302 \u00feK\u03d5xy \u2202\u00f0S\u03a6\u00de \u2202y\u0302 \u00feK \u2202 \u2202y\u0302 h\u0302 3 \u03d5yx \u2202\u00f0S\u03a6\u00de \u2202x\u0302 \u00feK\u03d5yy \u2202\u00f0S\u03a6\u00de \u2202y\u0302 \u00bc 6\u03be \u2202 \u2202x\u0302 1\u00fe\u00f01\u2212S\u00de\u03a6\u00bd \u00f0 ^hT\u2212S\u03d5c:c\u00de n o \u00fe6\u03beS \u2202\u03d5s:c:x \u2202x\u0302 \u00feK \u2202\u03d5s:c:y \u2202y\u0302 \u00f01\u00de In the liquid region, \u03a6\u22650 S\u00bc 1 p\u0302f \u00bc pf\u2212pcav pref\u2212pcav \u00bc S\u03a6 \u03c1\u0302\u00bc \u03c1 \u03c1f \u00bc 1 \u00f02\u00de In the cavitation region, \u03a6o0 S\u00bc 0 p\u0302f \u00bc pf\u2212pcav pref\u2212pcav \u00bc 0 \u03c1\u0302\u00bc \u03c1 \u03c1f \u00bc 1\u00fe\u00f01\u2212S\u00de\u03a6 \u00f03\u00de The boundary conditions are, \u03a6\u00bc p\u0302f \u00bc 0:1 at y\u0302\u00bc 0 (air-side) and y\u0302\u00bc 1 (oil-side)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002124_j.triboint.2010.05.005-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002124_j.triboint.2010.05.005-Figure2-1.png", + "caption": "Fig. 2. The ball-on-disc contact with spinning.", + "texts": [ + " Numerical calculations have also been completed to justify the experiment results and gain more insights into the film formation. Fig. 1 shows a schematic illustration of the optical EHL test rig used. The synchronous pulley A is driven by a serve-motor. The glass plate C rotates around its axis O1O2 at an angular speed o. The steel ball D is loaded against the glass plate C, and can rotate freely around its axis O3O4. The offset r of the Hertzian contact center with respect to axis O1O2 can be adjusted precisely by the XY table E. Fig. 2 presents the ball-on-disc configuration in more details. The velocity profile in the contact region along the y-axis is given in Fig. 3, which can be further decomposed into two components: pure rolling at the entrainment velocity ue and ion of the apparatus. spinning at an angular speed o. Velocities u1 and u2 at the Hertzian contact boundary on the disc are given by u1 \u00bco\u00f0r a\u00de u2 \u00bco\u00f0r\u00fea\u00de ( \u00f01\u00de where a is the Hertzian semi-contact width. And the decomposed velocities ur and us are expressed by ur \u00bc \u00f0u1\u00feu1\u00de=2\u00bcor \u00f02\u00de us \u00bc \u00f0u2 u1\u00de=2\u00bcoa \u00f03\u00de The steel ball\u2019s surface moves at a constant velocity of ur within the contact region", + " 9, it also appears that there is obvious decrease in the side lobe film thickness by a rise in load and the thickness difference between the side-lobes gets large. As seen from Eq. (4), when the semi-contact width a becomes large with increase in loads, the spin ratio increases, and therefore the fringes are more twisted. at different entrainment speed, w\u00bc16 N, Ssp\u00bc0.4. Fig. 10 gives the measured film thickness versus load at the two side-lobes, and hminR and hminL change in a similar way. In order to justify the observations in the experiments, full numerical calculations of EHL based on the model in Fig. 2 have been carried out in the present study. The lubricant is assumed to be Newtonian and only isothermal condition is applied. The numerical scheme is similar to that used by Yang and Cui [14] and will not be repeated here. Due to the limited rheological data available, the pressure\u2013viscosity coefficient of PB1300 was roughly estimated to be 3.0 10 8 Pa 1 from the experiments under pure rolling assuming that the pressure\u2013viscosity coefficient and the central film thickness obey the Hamrock\u2013Dowson formula" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000363_s11740-008-0109-1-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000363_s11740-008-0109-1-Figure4-1.png", + "caption": "Fig. 4 Experimental setup", + "texts": [ + " The whole degassing possibility is affected by thermal heat flow mechanisms in the liquid steel, in particular the marangoni convection [8], and interfered by the weld pool\u2019s surface tension. The mechanisms mentioned lead to unwanted pores within the weld seam, which can be seen in Fig. 3. 1.4 Experimental setup The experiments are carried out with a continuous wave 4 kW Yb:YAG laser, emitting at 1,030 nm. A focal spot diameter of 600 lm on the sheet\u2019s surface is used. The schematic drawing of the setup is depicted in Fig. 4. To monitor the optical process emissions, off- and onaxis photodiodes are applied in combination with a high speed video camera system. This offers the possibility to record image sequences with frame rates up to 25 kHz and to correlate them with the optical emissions. To visualize the weld pool surface, an external illumination by a diode laser emitting at 808 nm is selected. As welding material steel sheets and micro-alloyed sheets with a thickness of 1.25 mm up to 1.75 mm are used. Feed rates of 1", + " With the weld pool cooling down, the vapor tries to ascent through the liquid steel but often cannot ascent through the weld seam due to the surface tension, leading to failures and porosities. A schematic drawing of this effect is shown in Fig. 10. 1.7 External process influence A possible solution to decrease the monitored weld seam defects is to control the oscillations of weld pool and capillary by an external signal. Therefore the laser power is used as an input parameter and its amplitude is modulated with a frequency according to the eigenfrequency of the welding process. A system setup like in Fig. 4 can be used to detect the optical emissions, to calculate the modulation frequency and to communicate with the laser source. This is based on the assumption of an oscillating weld pool and keyhole, whose frequencies are both directly related to the laser power. At the eigenfrequency, the process transfer function will be maximized and its attenuation minimized. Hence, it is possible to force an adaptation to the external frequency, in order to prevent uncontrolled oscillations, resulting in an avoidance of a collapse" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000331_20070829-3-ru-4911.00048-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000331_20070829-3-ru-4911.00048-Figure10-1.png", + "caption": "Fig. 10. Trajectory of the aircraft.", + "texts": [ + " The purpose of the simulations is to verify the asymptotic tracking properties of the controller, both during standard demonstration manoeuvres (climb, steady turn, dive), and non-standard ones (such as a 180- degree roll). The flight plan includes the following sectors. (1) Initial speed 26 m/s. Climb from 30 m to 60 m at a rate of 12 m/s. (2) Accelerate to 42 m/s. (3) Perform a 180-degree turn of radius 200 m (at roll angle 43 deg). (4) Perform a 180-degree roll. (5) Perform a 180-degree turn of radius 200 m (at roll angle 137 deg). (6) Resume level flight. (7) Perform two successive 270-degree turns of radius 110 m (at roll angle 60 deg). The trajectory of the aircraft is shown in Figure 10. The aircraft follows the desired trajectory with zero tracking error for the airspeed and attitude and zero steady-state error for the altitude, despite the lack of information on the aerodynamics and despite the fact that we have used a simplified aerodynamic model for the design. Adaptive control of linear time invariant systems is a well\u2013established discipline whose major theoretical issues have been fully sorted out and the main plant structural obstacles clearly identified. Many successful applications of adaptive control for linear plants have also been reported in the literature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002113_09544062jmes2181-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002113_09544062jmes2181-Figure1-1.png", + "caption": "Fig. 1 Schematic of a rotary lip seal", + "texts": [ + " However, as the contact and fluid pressures are much lower than those in traditional elastohydrodynamic problems (metalto-metal contacts) and as the pressure dependence of the fluid viscosity does not play a very significant role, the lip seal problem is referred to as a problem in soft elastohydrodynamics. Two reviews of past studies on rotary lip seals exist. These cover both experimental and theoretical work on rotary seals [1, 2]. The present paper deals primarily with theoretical aspects and specifically with EHL modelling. Rotary lip seals come in a large variety of configurations. Most of these are described in reference [3]. The most common, and the most analysed, is shown in Fig. 1 and is frequently referred to as an \u2018oil seal\u2019. The lip is asymmetric, with a smaller angle between the lip and the shaft on the air side than on the liquid side. The region where the lip appears to meet the shaft is JMES2181 Proc. IMechE Vol. 224 Part C: J. Mechanical Engineering Science at WAYNE STATE UNIVERSITY on March 21, 2015pic.sagepub.comDownloaded from termed the sealing zone and is the primary focus of attention in an EHL analysis. Experimental studies have helped establish the basic principles of operation of these seals" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002343_s10846-011-9583-3-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002343_s10846-011-9583-3-Figure15-1.png", + "caption": "Fig. 15 The experimental setup", + "texts": [ + "1 Hardware Setup The experimental setup is composed of a DraganFlyer V Ti mini-rotorcraft, a four-channel GWS FM remote control radio unit, a Pentium IV PC with installed QuaRC 2.01 software for rapid prototyping from Quanser, and a 3-D tracker system (Polhemus Patriot), [33] used for measuring the position (X,Y,Z ) and orientation (\u03c6, \u03b8, \u03c8) of the quadrotor. The Polhemus Patriot system with an on-board position/orientation sensor is connected to the PC via RS 232 interface to allow feedback control (see Fig. 15). The PC generates control inputs, which are sent to the rotorcraft through the GWS radio control unit. The latter is connected to the computer via a Q4 Hardware-In-The-Loop board produced by Quanser. To simplify the experiments, each control input can be switched between automatic and manual control modes. Figure 16 shows the quad-rotor experimental platform in autonomous hover. The wires attached to the rotorcraft provide connections to the power supply and the position-attitude sensor. A wooden stick is attached to the rotorcraft for safety" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000031_6.2008-7413-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000031_6.2008-7413-Figure3-1.png", + "caption": "Figure 3. Side arms assembled (left) and side-arm rotor end (right).", + "texts": [ + " The body of the JAviator is formed like a cage, which is simple in its design, features fast and easy (re)assembly, and also ensures high integrity at relatively low weight. Compared to most other quadrotor bodies, this cage design acts like a protection frame for the onboard electronics and has proven in crashes to withstand collisions without serious damage. The four side arms consist of two CF pipes each with the same diameter as the vertical body pipes. Each side-arm pipe has AL connectors on both ends that fit to the AL flanges, as shown in Figure 3 (left), and also contain holes for mounting the rotor axles, shown in Figure 3 (right). In order to achieve a high degree of precision and reproducibility, the following fabrication methods were used. The CF rings and plates were manufactured by flow-jet cutting, whereas laser cutting was applied for the AL flanges, TI motor mounts, and TI rotor triangles. The AL frame connectors as well as all remaining AL and TI rotor components are CNC-fabricated. Due to the symmetrical frame design, the amount of 4 of 21 American Institute of Aeronautics and Astronautics comprising parts could be reduced to a lower number of different parts with higher quantity of occurrence in the following way" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.15-1.png", + "caption": "Fig. 2.15 Illustration of the Grasshopper escapement. a Front view b Back view", + "texts": [ + " Modified from the Anchor escapement, the Grasshopper escapement was used in his first three maritime time keepers: Harrison Number One (H1) through Harrison Number Three (H3). He then spent another 29 years on the project and finally won the Longitude Prize. Although the grasshopper escapement was not used in his final Harrison Number 4 (H4), which was a watch, it left a mark in history (Fig. 2.14). The Grasshopper escapement was also evolved from the anchor escapement. It has been suggested that the name of this escapement comes from the resemblance of the pallet arms to the legs of a grasshopper. As shown in Fig. 2.15, the escapement consists of an escape wheel, a pendulum, a driving mechanism (the lifted weight) and two pallets shaped like a grasshopper. Figure 2.16 gives the details of the escapement: The right pallet has an elbow joint connected to a heavy tail and a forearm, as well as a composer. The tail is slightly heavier so that the forearm tends to move away from the escape wheel. The composer prevents the forearm from rising further. As the upper arm rotates clockwise, the tip of the pallet at the forearm is pushed downwards" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001894_rnc.2813-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001894_rnc.2813-Figure1-1.png", + "caption": "Figure 1. Vertical take-off and landing aircraft.", + "texts": [ + " The paper is organized as follows. In Section 2, the control problem of the VTOL aircraft is formulated. In Section 3, the observer-based coordinate transformation is presented. In Section 4, the controller design procedure is proposed. In Section 5, the stability analysis is provided. In Section 6, the numerical simulation is given to show the effectiveness of the proposed design method. Section 7 draws the conclusions. On the basis of the nominal mathematical model of a VTOL aircraft [5], we consider the system shown in Figure 1, which takes input disturbances and unmatched uncertainties into account: Copyright \u00a9 2012 John Wiley & Sons, Ltd. Int. J. Robust. Nonlinear Control (2012) DOI: 10.1002/rnc Px1 D x2, Px2 D .u1C 1/ sin x5C .u2C 2/\" cos x5C f1.x, t /, Px3 D x4, Px4 D .u1C 1/ cos x5C .u2C 2/\" sin x5 1C f2.x, t /, Px5 D x6, Px6 D .u2C 2/C f3.x, t /, (1) where x1 and x3 are, respectively, the horizontal and vertical positions of the aircraft center of mass in the body-fixed reference frame; x5 is the roll angle of the aircraft; the control inputs are the thrust u1 (directed out the bottom of the aircraft) and the rolling moment about the aircraft center of mass u2; \" is a small coefficient that characterizes the coupling between the rolling moment and the lateral force; \u2018 1\u2019 denotes the normalized gravitational acceleration; 1 and 2 are thrust and rolling moment smooth disturbances, which are matched with the inputs, that is, the disturbances enter the system in the same channels as the inputs; the smooth functions fi " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003724_ccdc.2015.7162172-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003724_ccdc.2015.7162172-Figure1-1.png", + "caption": "Figure 1: A typical structure of quadrotor helicopter", + "texts": [], + "surrounding_texts": [ + "and military areas. Flight control is the foundation of all functions applied on quadrotor. PID controller, as a classical but reliable design, is widely used in quadrotor flight control. In the previous experiments, the PID coefficients are set empirically. In this paper, a coefficient tuning approach based on gradient optimization is introduced. The intended application is to tune the PID coefficients under an optimization cost function. This method is assessed by both simulation. The analysis of the experiment results gives suggestion in quadrotor design and parameter tuning.\nKey Words: Gradient Optimization, Quadrotor, PID tuning\n1 INTRODUCTION\nNowdays, micro-aerial vehicle (MAV) is becoming a hot topic. In military missions, MAV can be widely used in scouting, communication and transportation in the near future. In civilian aspect, some business-to-customer (B2C) have made attempts using MAV to do shipment in the last mile to customer. MAV, commonly in small size, have their own advantage compared with traditional unmanned aerial vehicles (UAVs), leading to an easy implementation in indoor environment. Also, the development in micro-electromechanical system (MEMS) and micro control unit (MCU) extend the range of MAV applications. Development of computation power make it possible to execute complex missions such as indoor exploration and mapping.\nQuadrotor\u2019s first concept came from 1907, very near the birth of the Wright brothers\u2019 plane. With the development mechanics and electronics in the last century, it becomes very popular in MAV design these years. Because of its simple structure, it is very easy to build and control, and also reliable in mechanical aspect. The use of four propellers makes its size smaller than traditional single-propeller helicopter under same take-off weight. Compared to complicated mechanical structure of helicopter, quadrotor changes flight direction simply by changing the rotation speeds of four motors.\nBy now, various controllers have been implemented on quadrotor [1, 2, 3]. The common controllers are old but reliable PID, because it is suitable for highly nonlinear and coupled dynamic systems. Some literatures [4, 5] introduced linearized model around equilibrium point. But it now always work under the condition of drastic maneu-\nThis work is supported by National High Technology Research and Development Program of China (863 Program) (2012AA041701), National Natural Science Foundation of China (61104048, 61473253). Corresponding author: Chao Xu, cxu@zju.edu.cn\nvering. After the dynamic model of quadrotor is well researched [6], the feedback gain matrix is derived [5] from state space system. Since accurate model is hard to derive [7, 8] for each quadrotor, it is not commonly used. The update of PID controller is cascade PID controller [9]. The inner loop stabilizes the angular rates, which are directly measured by gyroscope. The outer loop is in charge of stabilizing Euler angles. Euler angles are derived from data fusion and filtering of gyroscope and accelerometer measurements. Since the problem of linear programming first established in the 1930s, optimization has become one of the most important topics in control theory. Optimal control uses optimization method to derive the best control strategy under some metric [10]. The major part of cost function is in quadratic formation, which is a special case of convex optimization problem. Gradient descent method is the most widely used approach in quadratic programming (QP) problems. For nonlinear dynamic systems, gradient descent method is not enough because there is no direct way to derive gradient in a dynamic system. Variational system\n1588978-1-4799-7016-2/15/$31.00 c\u00a92015 IEEE", + "is established due to this problem. Loxton and Lin [11, 12, 13] introduced dynamic gradient method based on variational system, which can well address this kind of dynamic optimization problem. In the practical experiments, the most important part in designing a flight controller is tuning the PID coefficients. Even under same control structure, different coefficients lead to different performance. Usually, the PID coefficients are set by experiments and experience to the best of our knowledge. Few research is established in quadrotor PID tuning area. This paper provides a tuning approach based on optimization method through variational system. This method gives tuning results for quadrotor\u2019s nonlinear, highly-coupled system, and shows good performance in simulation. The rest of this paper is organized as follow, section 2 will introduce the dynamic model of quadrotor briefly. Section 3 will demonstrate the gradient optimization method based on variational system. Section 4 will show how to adapt gradient optimization method on quadrotor model. Section 5 will analyze the simulation result and section 6 will draw a conclusion.\n2 Quadrotor Dynamics\nQuadrotor is a different kind of aerial vehicle compared to a traditional helicopter in the manner of dynamics and control. But states which describe the attitude of a aerial vehicle remain the same: the velocity corresponding to body axis (u, v, w), the angular rate toward body axis (p, q, r), the position toward earth axis (x, y, z), the Euler angle roll, pitch, yaw (\u03c6, \u03b8, \u03c8). As the sensors onboard are not sufficient for all information, only 6 states (p, q, r, \u03c6, \u03b8, \u03c8) are observable without external support. Though the dynamics of quadrotor is relatively simpler than helicopter, the analysis to determine the whole dynamics model is in great length. The way deriving this dynamic model will not be discussed in this paper. Only a brief conclusion sufficient to support section 4 will be given. The index of coordinates of quadrotor are defined in figure 2. There are two frames used in this dynamic analysis, the earth frame E and the body frame B. The earth frame is fixed to the earth ground. And the body frame is on the aircraft itself. In the earth frame, the resultant force of lift force and gravity decide the motion on z-axis. The kinematics and dynamics on z-axis is as:{ z\u0307 = w w\u0307 = (U1 \u00b7 cos\u03c6 \u00b7 cos \u03b8 \u2212m \u00b7 g)/m . (1)\nThe relation between earth frame can be represented by a transformation. Here the details of transformation will not be discussed. The derivative of Euler angle is a function of Euler angle \u03c6, \u03b8, \u03c8 and angular rate p, q, r. \u23a7\u23a8 \u23a9 \u03c6\u0307 = (p \u00b7 cos\u03b8 + q \u00b7 sin\u03c6 \u00b7 sin\u03b8 + r \u00b7 cos\u03c6 \u00b7 sin\u03b8)/cos\u03b8 \u03b8\u0307 = q \u00b7 cos\u03c6+ r \u00b7 sin\u03c6 \u03c8\u0307 = (q \u00b7 sin\u03c6+ r \u00b7 cos\u03c6)/cos\u03b8 .\n(2)\nFor the derivative of angular rate, there are three parts taken into consideration: the torque generated by propeller\u2019s lift, the angular momentum of body, and gyroscopic effect of propellers. Combine all three parts together, and the differentiate equation to the angular rate is: \u23a7\u23a8 \u23a9 p\u0307 = [ \u221a 2 \u00b7 l \u00b7 U2 + q \u00b7 r \u00b7 (Iy \u2212 Iz)\u2212 JTP \u00b7 q \u00b7 \u03a9]/Ix q\u0307 = [ \u221a 2 \u00b7 l \u00b7 U3 + p \u00b7 r \u00b7 (Iz \u2212 Ix)\u2212 JTP \u00b7 p \u00b7 \u03a9]/Iy\nr\u0307 = [ \u221a 2 \u00b7 U4 + q \u00b7 p \u00b7 (Ix \u2212 Iy)]/Iz\n.\n(3)\nNote that the gyroscopic effects of propellers, as JTP \u00b7p \u00b7\u03a9 and JTP \u00b7 q \u00b7 \u03a9 in equation (3), are very small compared with other entries. For computational convenience, they are eliminated in the optimization problem in section 4.\nIn the equation sets above, U1, U2, U3, U4 are the controller outputs of four feedback loops for z, \u03c6, \u03b8 and \u03c8. They are manipulated by the rotation rate change of four propellers. The control outputs are decoupled to \u03c9i using the equations below in real experiments. But \u03c9i will not be further used in the rest of this paper. Ui will remain separately in the format of PID as described in the coming section 4.\n\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 U1 = pt \u00b7 \u22114 i=1 \u03c9 2 i\nU2 = pt \u00b7 (\u03c923 + \u03c924 \u2212 \u03c921 \u2212 \u03c922) U3 = pt \u00b7 (\u03c921 + \u03c922 \u2212 \u03c922 \u2212 \u03c924) U4 = \u03bc \u00b7 pd \u00b7 (\u03c921 + \u03c924 \u2212 \u03c922 \u2212 \u03c923) . (4)\nFor all the control output, single loop PID is used for 4 states, z, \u03c6, \u03b8, \u03c8. Conventionally, simple PD controller is enough to keep a quadrotor hovering. But in purpose of eliminating the steady-state error in height control, integral part is add to the controller, which make the design of gradient method far more complicated than simple PD controller. This part will be discussed in section 4.\nTable1 Variable description\nParameter Unit Descritpion\nz m height w m \u00b7 s\u22121 velocity on height \u03c6 rad roll angle \u03b8 rad pitch angle \u03c8 rad yaw angle p rad \u00b7 s\u22121 angular rate on x-axis q rad \u00b7 s\u22121 angular rate on y-axis r rad \u00b7 s\u22121 angular rate on z-axis U1 N control output on height U2 N control output on roll U3 N control output on pitch U4 N \u00b7m control output on yaw\n2015 27th Chinese Control and Decision Conference (CCDC) 1589" + ] + }, + { + "image_filename": "designv11_3_0002865_1.3555006-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002865_1.3555006-Figure11-1.png", + "caption": "Fig. 11 Fig. 12", + "texts": [ + " Tests were also carried out under a suddenly applied load representing practically a step-function loading. In middle of 1-in. dia. Tests were carried out on this bearing under the conditions shown in Table 1. In tests Nos. 1 and 2, the same type of restrictor was used, but with different values of film thickness h and hence different a and (3CJO. Test No. 3, however, was similar to No. 2 regarding the values of a and /3co0, but with a different type of restrictor for the purpose of comparison. 2 A scale-model machine table, Fig. 11, supported on four all tests the fluid used was \"Shell-Tellus 27\" oil. Diaphragm Controlled Restrictor (DCR).2 The DCR [3] is a variable hydraulic restrictor, the resistance of which is controlled by the deflection of a diaphragm as a result of the feedback bearmg recess pressure. The fluid is supplied at constant pressure through the inlet port A, Fig. 12. I t then passes through the controlling pad (1) to the bearing via the outlet port B. The bearing recess pressure (pb) is always acting on the diaphragm (2) which forms one side of the controlling pad", + " The movement of the bearing was measured by a relative velocity pickup (P) which was fixed to the lower cylindrical block (3) with its probe in contact with the upper part of the bearing (5). The pressure in the cylinder as well as the bearing recess pressure was measured by means of two electrohydraulic pressure transducers. Signals from the velocity pickup and the pressure transducers were amplified and then fed to the galvanometers of an ultraviolet recorder, Fig. 10(a). Scale-Model Table Test Rig. The arrangement used for testing the scale-model machine table is shown in Fig. 11. Two rails, integral with the bed (1), had eight bearings\u2014four to carry the vertical load and the other four to carry the horizontal load. Two slideways in the slider (2) formed the upper part of the bearings. This test rig was designed in such a manner as to eliminate the effect of any mechanical friction. Hence, the initial static vertical load as well as the step load was produced by means of a loading pad (3) fixed to the middle of the table. Three of the upper bearings were connected to three diaphragm controlled restrictors, while the fourth was compensated by a capillary restrictor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003222_s10626-013-0169-z-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003222_s10626-013-0169-z-Figure3-1.png", + "caption": "Fig. 3 The history of \u2016xk1\u2016 and xk2 with events (35)", + "texts": [ + " Assume that sensor 1 can sample x1 := [z1, z2]T and sensor 2 samples x2 := z3. Also assume that there is a constant transmission delay of 0.01s in the communication channel. In the first experiment, the regular state-dependent event (8) is used. We set \u03c11 = \u03c12 = 0.05 for both agents, the events are defined as Agent 1: \u22120.05\u2016xk1\u2016 + \u2016x1(t)\u2212 xk1\u2016 = 0 Agent 2: \u22120.05\u2016xk2\u2016 + \u2016x2(t)\u2212 xk2\u2016 = 0. (35) The event-triggered NCS is simulated in MATLAB for 40 s. The state trajectories are plotted in Fig. 2. We can see that the event-triggered scheme stabilized the system. Figure 3 depicts the trajectories of \u2016xk1\u2016 and xk2 with broadcast events (star) marked on the trajectories. Figure 4 provides a close-up look revealing more details. The time between two consecutive events are presented in Fig. 5. It can be observed that when \u2016x1(t)\u2016 approaches 0, the frequency of events increases. In the simulation when xk2 crosses 0, since we cannot have arbitrarily small step size, we do not exactly reproduce Zeno triggering phenomenon. But a drastic increase of triggered events can be observed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure8.6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure8.6-1.png", + "caption": "Fig. 8.6 Scheme of the mechanism with spring", + "texts": [ + " Hercog BookID 175907_ChapID 8_Proof# 1 - 14/4/2009 BookID 175907_ChapID 8_Proof# 1 - 14/4/2009 2 , 2 = c c c m r J (8.2) and the inertia of stainless steel disc with mass m s and radius r s is: 2 . 2 = s s s m r J (8.3) The aluminium part at the back of the mechanism can be modelled as a point mass with inertia: 2 .=a aJ m r (8.4) For the ingot at the back the equation for inertia is: = 2 / 12.p p pJ m l (8.5) The total inertia is the sum of all four inertias: .c s a pJ J J J J= + + + (8.6) For the derivation of the spring torque, the scheme of the spring mechanism shown in Fig. 8.6 will be used. The spring torque sT r is the cross product of the position vector r r and the spring force sF r : 8 Teaching of Robot Control with Remote Experiments 179 BookID 175907_ChapID 8_Proof# 1 - 14/4/2009 replaced by the sinus of the angle between position vector r r and force sF r : ( ) ( ).s s sT rF sin rF sinp a a= \u2212 = \u2212 (8.8) The spring force sF r is the product of the spring extension D x and the coefficient of elasticity k (Hook\u2019s law): .sF k x= \u2206 (8.9) The angle a from Eq. (8.8) and extension of spring D x from Eq. (8.9) must be expressed as the functions of two known dimensions r, l and measured angle q. In Fig. 8.6 one can see that the two sides of the triangle, and the angle q between them are known. The first unknown value is the angle a or even better sin (a), which is directly required in equation Eq. (8.8). Also, the spring extension D x has to be calculated. It cannot be calculated directly but only from the third side of the triangle, that is the length d of the extended spring. By using the Law of Cosines, the following expression for the length of the extended spring, d, can be written: ( ) ( ) ( )22 2 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002068_iicpe.2012.6450421-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002068_iicpe.2012.6450421-Figure1-1.png", + "caption": "Fig. 1. Stator and rotor flux linkage in coordinate system", + "texts": [ + " abc model, two axis dq-model have been proposed for different applications, the two axis dq-model is simple and is widely used. The dynamic model of PMSM is derived from two phase synchronous (stator) reference frame. For dynamic model of PMSM, the assumptions made are - spatial distribution of magnetic flux in air gap should be sinusoidal, and magnetic circuit should be linear (hysteresis and eddy current losses are negligible) [1, 4, 13]. The stator flux reference frame in D axis is in phase with stator flux linkage space vector 'Ps. Q axis (of SRF) leads 90\u00b0 to D axis in CCW direction as depicted in Fig 1. (1) Where, 9s= rotational angle of stator flux vector 9r=rotational electric angle of rotor 9s =9r+o Stator flux linkage is given by (2) Where Ls is stator self inductance and 'Paf is the rotor permanent magnet flux linkage. The stator voltage equation in rotor reference frame (dq reference frame) are given as . d\\Pd Vd = Rsld + dt - wr\\Pq (3) (4) Where 'I' q= Lqiq and 'I'd = Ldid+'I'af, Vd and Vq are d-q axis stator voltages, id and iq are d-q axis stator currents, Ld and Lq are d-q axis inductances" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002909_s12206-012-0312-z-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002909_s12206-012-0312-z-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of the physical domain involving an elastic rod surrounded by a viscous fluid.", + "texts": [ + " The existence of a critical value of angular frequency of rotation of the motor which demarcates twirling and overwhirling motion is also investigated. The paper is arranged as follows: The elastic rod model based on a network of various elastic links and fluid model based on the momentum and continuity equations and the numerical procedure are explained in Section 2. The numerical results generated by employing it are discussed in Section 3. Finally, the concluding remarks of the present study appear in Section 4. Fig. 1 shows the schematic diagram of the physical domain which involves an elastic and intertialess rod surrounded by a viscous fluid at a very low Reynolds number. The rod is modelled as a flexible cylindrical one with a motor attached to its base at one of the ends. The other end is free and kept slightly bent. The rod is divided into a number of circular rings and each ring has twelve IB points. The first circular cross-section is modelled as the motor part. Rotation of the motor and hence the whole rod is generated by an applied angular frequency of rotation of the motor part", + "XUn n n ib k k h\u03b4+ += \u2212\u2211 (15) The IB point 1 2,Xn k k is then moved to its new position at this Lagrangian velocity according to the following equation 1 2 1 2 1 , , 1= +X X Un n k k k k n ibt+ +\u0394 (16) where t\u0394 is the time-step size. The governing equations outlined in Section 2 are solved in a three\u2013dimensional dimensionless cubic domain of 20 20 20\u00d7 \u00d7 with periodic boundary conditions in all the three directions. We use 32 32 32\u00d7 \u00d7 grid points with uniform Eulerian grid size. Here the numerical simulations are performed for an elastic rod of length, 10.0l = , and diameter, 1.0d = , as shown in Fig. 1. The initial position of the rod is set at 5.0, 10.0, 10.0c c cx y z= = = , where ( ), ,c c cx y z are the coordinates of the center of the first circular cross-section. We use 41 circular cross-sections along the axial length of the rod. The stiffness constants of all the elastic links are set as 1.0c l dS S S= = = , where ,c lS S and dS are the stiffness constants of cross-sectional, longitudinal and diagonal links, respectively. The fluid is at rest at the beginning of the simulation. A numerical model is developed to simulate the present problem based on an immersed-boundary finite-volume method", + " We will present all the numerical results in a dimensionless form and the conversion to the dimensional form can be easily obtained using the viscosity of water, 0.001 / ,kg ms\u03bc = characteristic length(rod diameter), 510 ,cl m\u2212= and characteristic velocity, 55 10 / .cU m s\u2212= \u00d7 In this work, the behavior of the rod for a range of values of angular frequency of rotation of the motor is examined. In this section, we present the numerical simulation results relevant to the behavior of the elastic rod represented in Fig. 1 when driven by an applied angular frequency of the motor 0.001\u03c9 = in a viscous fluid. Fig. 4 depicts the two\u2013 dimensional view of the instantaneous shape of the rod during the motion. It can be seen that a straight and twisted rod rotates about its axis for this low value of rotational frequency. This motion is referred to as twirling motion. The rod attains a stable straight state at the end of the simulation similar to its equilibrium state. The applied angular frequency induces rotation at the motor part of the rod in the counter-clockwise direction when viewed from the motor end of the rod and the various elastic links become activated due to the change in their resting lengths, which induces necessary elastic forces and results into the rotation of the entire rod", + " 5(a)-(d) illustrates the instantaneous fluid velocity vector plot at the last (non-motor end) cross-section of the rod when viewed from the front (motor end) side of the rod, together with the rod. A rotating flow field is observed around the rod along its length at all the times. It is evident from the figure that the rotating flow direction is counter-clockwise when viewed from the motor end of the rod. The numerical simulations are performed to study the dynamic behavior of the elastic rod depicted in Fig. 1 by applying a high value of angular frequency of rotation of the motor 0.004\u03c9 = . Fig. 6 illustrates the two\u2013dimensional view of the instantaneous shape of the rod during the motion. From the figure it can be inferred that the rod initially undergoes whirl ing motion in which the rod takes a helical shape and later a discontinuous shape transition occurs resulting into the phenomenon of overwhirling. In the overwhirling motion the rod almost folds back on itself. Also, the free end of the rod goes below the motor end at the end of the simulation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001544_tmag.2010.2064331-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001544_tmag.2010.2064331-Figure2-1.png", + "caption": "Fig. 2. Coordinate transformations for the shifted coil in free space. A denotes the center of the coil and O is the origin of the unprimed coordinate system.", + "texts": [ + " 1 reads [19] (22) (23) Substituting in the expressions (10) and (11) and using again the orthogonality of the exponential functions, the continuity relations (19)\u2013(21) give after some manipulations (24) (25) (26) The expression for the excitation term of an air-cored bobbin coil is derived by the integral representation of the azimuthal component of the magnetic vector potential produced in freespace [12] (27) where (28) and the position vector being defined with respect to the center of the coil (cf. Fig. 2). In (28), stand for the inner and outer radii of the coil, is its height, and denotes the wire turns. The corresponding expression for is (29) Applying the addition theorem to the triangle OAB (cf. Fig. 2), the Bessel function can be written in analogy with (23) (30) where and have been exchanged with and , respectively, and is the angle between and . Substituting (30) into (29) and comparing it with (31) it turns out that (32) In this work, we focus on the bobbin coil. However, any other coil shape or orientation can be modeled as long as the excitation term can be computed. Another coil excitation of interest in eddy-current tube testing is the horizontal coil, i.e., a pancake type coil with axis that is perpendicular to the inner tube surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003139_j.triboint.2013.10.010-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003139_j.triboint.2013.10.010-Figure1-1.png", + "caption": "Fig. 1. Physical configuration of a journal bearing system lubricated with a nonNewtonian couple stress fluid.", + "texts": [ + " Although the running speed below the stability threshold speed yields linearly stable behavior according to the linearized stability theory, unstable orbits for the non-Newtonian fluid lubricated bearing depending upon the initial positions are illustrated through the transient nonlinear analysis. In addition, the influences of non-Newtonian fluids on the size of the nonlinear stability boundary within the clearance circle are investigated through the variation of the non-Newtonian couple stress parameter. Fig. 1 describes the physical configuration of a journal bearing system. The journal of R is rotating with an angular velocity \u03c9n within the bearing housing. The film thickness is hn \u00bc C\u00fee cos \u03b8, where C is the maximum clearance, e is the eccentricity, \u03b8\u00bc xn=R is the circumferential coordinate. The lubricant in the film region is taken to be a non-Newtonian incompressible couple stress fluid of Stokes [12]. Under the usual assumptions of thin-film lubrication theory, the non-Newtonian dynamic Reynolds equation governing the dynamic film pressure pn can be expressed as Lin [21]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003234_1.4003996-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003234_1.4003996-Figure5-1.png", + "caption": "Fig. 5 Plastic strain at dimensionless time t=tc = 0.22 corresponding to maximum sphere penetration (Pemax = 0.32, Vy=Vy_c = 157.3, Ef=Yf = 106, m = 0.33, and l = 0.3)", + "texts": [ + " The dimensionless contact area shown in Fig. 3 between the sphere and the flat increases, reaches a maximum, and decreases thereafter. The maximum of the contact area occurs at the maximum penetration of the sphere into the flat. Figure 4 shows the dimensionless vertical displacements as a function of the dimensionless time for a point fixed at the bottom of the sphere and a point fixed on the surface of the flat. The location of the point on the flat surface corresponds to the maximum penetration of the sphere as shown in Fig. 5. The dimensionless vertical displacement of the bottom of the sphere, marked by the dashed curve in Fig. 4, follows the change in the dimensionless vertical velocity shown in Fig. 3, i.e., the sphere penetrates up to the maximum depth of y=xc = 65 around t=tc = 0.22 and then returns back to zero after complete separation at t=tc = 0.42. When the fixed point on the moving flat is far from the impacting sphere, its displacement is zero up to a dimensionless time of approximately 0.1. Journal of Tribology JULY 2011, Vol. 133 / 031404-3 Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 08/24/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use As the fixed point approaches the impacting sphere, it deflects below the original surface profile (y=xc = 0) and then rapidly reaches its maximum value as a result of pile-up at the leading edge of the contact zone (see Fig. 5). Later on, when the bottom of the sphere passes the fixed point on the moving flat, it deflects below the original flat surface profile and the dimensionless vertical displacement value coincides with the sphere maximum penetration. After the fixed point on the moving flat has passed the sphere completely, its vertical displacement recovers due to elastic recovery of the flat. It can be seen from Fig. 4 that the final dimensionless deflection of the fixed point on the flat is 54. This value corresponds to the residual maximum penetration depth on the flat surface xres_max. Figure 5 shows the plastic strain in the contact zone at the dimensionless time t=tc = 0.22 corresponding to the maximum penetration. We observe that pile-up occurs at the leading edge of the contact due to plastic deformation of the contact zone. Figure 6 shows the evolution of the dimensionless contact force and the dimensionless friction force during a contact. The friction force Q is composed of two terms. The first term corresponds to pure sliding friction at the contact interface. The second term is plowing friction due to the penetration of the sphere into the flat", + " During the rebounding of the sphere, the penetration depth and the total friction coefficient decrease. The final value of the friction coefficient equals 0.3 corresponding to pure sliding friction. The mean contact pressure p, defined as the ratio of the contact force P and the contact area A normalized by the yield strength Y of the soft flat, is shown by the solid curve in Fig. 7. If the normalized mean contact pressure exceeds 1.07, plastic deformation occurs in the contact zone [30] as can also be seen in Fig. 5. Figure 8 shows the evolution of the dimensionless temperature rise for the two fixed points shown in Fig. 5. We observe that the location of the maximum temperature in the contact zone changes during a transient contact. The temperature is given in Figs. 9(a) and 9(b) for two dimensionless contact times of 0.03 and 0.29, respectively. At the beginning of a contact, the location of the maximum temperature is in the middle of the contact zone as 031404-4 / Vol. 133, JULY 2011 Transactions of the ASME Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 08/24/2017 Terms of Use: http://www" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000008_1.2538787-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000008_1.2538787-Figure2-1.png", + "caption": "FIG. 2. a Experimental setup for inscription of surface relief gratings in azo thin films. The azo sample S , cast on a glass slide is placed in proximity to a mirror M . An inscription laser beam Ar+ 488 nm is circularly polarized using a quarter-wave plate WP . The laser beam reflects off of the mirror and interferes with itself at an angle . The grating formation can be measured in real time using a probe laser beam HeNe 633 nm that is diffracted into a detector D . b The resulting sinusoidal surface relief grating, inscribed into the azo-polymer surface, as measured using atomic force microscopy.", + "texts": [ + " This fully reversible chromophore alignment leads to strong dichroism and birefringence in the azo materials due to the azo\u2019s anisotropic structure. The azo photomotion has been investigated as a photoswitch,3 to align liquid-crystalline systems,4 to bend freestanding thin films,5 and in many other photodynamic systems.2 In 1995, a further photomotion was discovered in the azo system.6 It was found that irradiation with a light intensity and/or polarization gradient would lead to spontaneous large-scale motion of polymer material in the thin film, resulting in surface topographical patterning. In the simplest experiment see Fig. 2 , the sample is exposed to a sinusoidal variation in light intensity, originating from the interference of two coherent laser beams. The end result is that the azo material generates a surface relief grating SRG , as shown in Fig. 2. Any incident light field can be used, with the azo material\u2019s surface deforming to reproduce the light pattern. The process is strongly polarization dependent7 and occurs even at remarkably low laser power 1\u2013100 mW/cm2 . This topographical hologram is stable indefinitely at room temperature, but can be erased by heating the film past its glassto-rubber transition temperature Tg , in which case the original film thickness is recovered. This demonstrates that the process in question is a material motion, and not a photoablation phenomenon where material is lost, as would be the case for most materials irradiated at extremely high power", + " Since the final film thickness depends on the drag velocity as well as solution concentration , an acceleration ramp will generate a thin film with a thickness gradient. In this work, a 50 l drop of polymer solution 5 10\u22122 mol/ l, based on repeat unit, pdr1a in THF was coated onto cleaned glass microscope slides. The gap between the substrate and the blade was maintained at 100 m, with a blade angle of 5\u00b0. The substrate was ramped from 1 to 5 mm/s at an acceleration of 0.56 mm/s2. SRGs were prepared by placing a thin film adjacent to a mirror, which reflected a coherent laser beam onto the surface experimental setup shown in Fig. 2 . Inscription was performed at room temperature 25 \u00b0C , and from previous work it is known that the low laser power used does not lead to appreciable sample heating.21 Laser inscription was performed using the 488 nm line of an argon-ion laser Coherent Innova 308 . Unless otherwise specified, the irradiation intensity was 37 mW/cm2, the inscription angle was =20\u00b0, and the irradiation time was 420 s. The incident laser beam was circularly polarized using a quarter-wave plate. Interference of right- and left-handed circularly polarized laser beams is known to give rise to high-efficiency surface relief formation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003140_1.4005032-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003140_1.4005032-Figure4-1.png", + "caption": "Fig. 4 Interaction between two gears in the process of back side collision status", + "texts": [ + "org/about-asme/terms-of-use Wpx \u00bc Fm\u00f0t\u00de sin a\u00fe mpepx2 p cos\u00f0xpt\u00de Wpy \u00bc Fm\u00f0t\u00de cos a\u00fe mpepx2 p sin\u00f0xpt\u00de Mp \u00bc Tp Fm\u00f0t\u00derp Wgx \u00bc Fm\u00f0t\u00de sin a\u00fe mgegx2 g cos\u00f0xgt\u00de Wgy \u00bc Fm\u00f0t\u00de cos a\u00fe mgegx2 g sin\u00f0xgt\u00de Mg \u00bc Fm\u00f0t\u00derg Tg 8>>>>>>< >>>>>>: (8) Wpx \u00bc k\u00f0t\u00def \u00f0D\u00f0t\u00de\u00de sin a\u00fe mpepx2 p cos\u00f0xpt\u00de cm\u00f0\u00f0 _xp _xg\u00de sin a\u00fe \u00f0 _yp _yg\u00de cos a\u00fe rp _hp rg _hg _e\u00f0t\u00de\u00de sin a Wpy \u00bc k\u00f0t\u00def \u00f0D\u00f0t\u00de\u00de cos a\u00fe mpepx2 p sin\u00f0xpt\u00de cm\u00f0\u00f0 _xp _xg\u00de sin a\u00fe \u00f0 _yp _yg\u00de cos a\u00fe rp _hp rg _hg _e\u00f0t\u00de\u00de cos a Mp \u00bc k\u00f0t\u00derp\u00f0\u00f0xp xg\u00de sin a\u00fe \u00f0yp yg\u00de cos a\u00fe rphp rghg e\u00f0t\u00de\u00de cmrp\u00f0\u00f0 _xp _xg\u00de sin a\u00fe \u00f0 _yp _yg\u00de cos a \u00fe rp _hp rg _hg _e\u00f0t\u00de\u00de \u00fe Tp Wgx \u00bc k\u00f0t\u00def \u00f0D\u00f0t\u00de\u00de sin a\u00fe mgegx2 g cos\u00f0xgt\u00de \u00fe cm\u00f0\u00f0 _xp _xg\u00de sin a\u00fe \u00f0 _yp _yg\u00de cos a\u00fe rp _hp rg _hg _e\u00f0t\u00de\u00de sin a Wgy \u00bc k\u00f0t\u00def \u00f0D\u00f0t\u00de\u00de cos a\u00fe mgegx2 g sin\u00f0xgt\u00de \u00fe cm\u00f0\u00f0 _xp _xg\u00de sin a\u00fe \u00f0 _yp _yg\u00de cos a\u00fe rp _hp rg _hg _e\u00f0t\u00de\u00de cos a Mg \u00bc rgk\u00f0t\u00de\u00f0\u00f0xp xg\u00de sin a\u00fe \u00f0yp yg\u00de cos a\u00fe rphp rghg e\u00f0t\u00de\u00de \u00fe rgcm\u00f0\u00f0 _xp _xg\u00de sin a\u00fe \u00f0 _yp _yg\u00de cos a\u00fe rp _hp rg _hg _e\u00f0t\u00de\u00de Tg 8>>>>>>>>>>>>>>>>>>>>>>>>>>>>>>< >>>>>>>>>>>>>>>>>>>>>>>>>>>>>: (9) Here, Wix and Wiy(i \u00bc p; g) are the excitations acting on the driving gear and driven gear, and Mi(i \u00bc p; g) is the moment excitation acting on the gears. When the meshing status becomes back side collision, then the direction of dynamic mesh force will change as Fig. 4 shows. The dynamic force Fm will become negative, and the dynamic force which acts on the driving and driven gear can be defined as Wpx \u00bc Fm\u00f0t\u00de sin a\u00fe mpepx2 p cos\u00f0xpt\u00de Wpy \u00bc Fm\u00f0t\u00de cos a\u00fe mpepx2 p sin\u00f0xpt\u00de Mp \u00bc Tp Fm\u00f0t\u00derp Wgx \u00bc Fm\u00f0t\u00de sin a\u00fe mgegx2 g cos\u00f0xgt\u00de Wgy \u00bc Fm\u00f0t\u00de cos a\u00fe mgegx2 g sin\u00f0xgt\u00de Mg \u00bc Fm\u00f0t\u00derg Tg 8>>>< >>>: (10) 2.2 Nonlinear Oil Film Force. The dimensionless Reynolds equation [16] is defined as R L 2 @ @z h3 @p @z \u00bc x sin# y cos# 2\u00f0 _x cos#\u00fe _y sin#\u00de (11) The dimensionless oil pressure can be gained by using Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003412_j.issn.1004-4132.2011.04.017-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003412_j.issn.1004-4132.2011.04.017-Figure5-1.png", + "caption": "Fig. 5 Topology 2", + "texts": [ + "3, we find that each agent converges to the leader\u2019s states with the initial states of the agents generated randomly and the static leader\u2019s position chosen as x0 = 3 (see Fig. 3). However, we find that the LMI (9) has no feasible solutions for the large enough control gain, i.e., the control gain k can not be too large. Example 2 Switching interconnection topology. Consider a network of four agents and a static leader given by (10). The interconnection topology of the system is switched between the topology 1 in Fig. 4 and the topology 2 in Fig. 5. we choose \u03b31 = 0.03, \u03b32 = 0, \u03b33 = 0.02, \u03b34 = 0.01, and take k1 = k2 = k3 = k4 = 4 for simplicity. Using the LMI toolbox in Matlab for the LMI (13) in Theorem 3, the maximum communication delay is \u03c4max = 2.018 s, i.e., the LMI (13) holds for each topology if \u03c4 2.018 s. In the simulation, the interconnection topology is switched from one topology to another one every 3 s. With the initial states of the agents generated randomly, the static leader\u2019s position is chosen as x0 =3. Under the communication delay \u03c4 = 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002516_ed044p172-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002516_ed044p172-Figure1-1.png", + "caption": "Figure 1. Hypothetical reaction progre.3 curves illurtroting principles of the methods.", + "texts": [ + " The increase of absorbance with time, which represents the decrease in glucose concentration, provides sufficient kinetic data to ascertain the initial glucose concentration by a modification of the single-point method (1-A) of Worthington Biochemical Corp.' and by modification of the initial rate method of Malrnstadt and Both methods were used simultaneously on each standard and unknown sample. Recent work by Malmstadt and Croucha describes an automated experiment in which glucose is determined by coupling the production of H202 to the iodine-iodide couple. Figure 1, hypothetical reaction progress curves for two different initial glucose concentrations, demonstrates the orincioles of the two methods and denotes horeeradiah the symholsused h this discussion. Ha01 + O-dianisidine _;--t H1O + o-dianisidine (oxidized) For the conditions of this experiment, the rate of dis- perar1usse appearance of glucose may be formulated from the For ease in following the discussion, the above reac- Michaelis-Menton model: tions may be represented as: dG - k'.G.E.02 enaymes -- dt - K, + G (1) glucose + 0 1 + o-dianisidine ---+ buffer gluconic acid + o", + " G dl (2) 172 / Journal o f Chemical Education where k is a pseudo-first-order rate constant depending on E, 07, temperature, and pH. Eqn. (2) may be integrated to obtain: G = Gee\" (3) where Go is the initial glucose concentration. Variable-Time Method The variable-time method is described by Blaedel and hick^.^ The instantaneous glucose concentration a t an arbitrary absorbance A is related to the initial glucose concentration by: GO - G = A/ab (4) where a is the absorptivity of the oxidized form of o-dianisidine at 440 m, and b is the effective cell pathlength. At arbitrary but constant values of AI and A, as shown in Figure 1, the instantaneous concentrations GI and Gz may be related through eqn. (4). GI = Go - Atlab (5) Gz = GO - Adah Gz is related to GI by eqn. (3) as follows: Gz = G,e-[h(lz-fi)l = GIe-kAr (6) Therefore, substitution of eqn. (5) into (6) yields: which exactly expresses the relationship of Go to the experimentally measured time interval, At = h - tl, ' MALMSTADT, H. Ti., AND HICKS, G. P., o p . cd. 4 BLAEDEL, W. J., AND HICKS, G. P. (Editor: Reilley, C. N.), \"Advances in Analytical Chemistry and Instrumentation,\" Vol", + " 8 Prepared enzymatic glucose reagent, Worthington Biochemical Corporation, Freehold, N. J. 'See footnote 8. and predetermined absorbance values A1 and A2. Equation (7) is simplified by expanding the exponential and retaining only the first two terms. Go - A8/ab 2. (Go - A,/ah) (1 - kAt) (8) Equation (8) on rearrangement predicts a linear calibration curve of Go versus l/At with intercept A,/ab which corresponds to the amount of glucose rearted to obtain Al. The experimental verification of the form of eqn. (9) is shown in Figure 2A. Figure 1 illustrates the inverse dependence of At on Go. For Goa > G,\" it is seen that At, < At,. It is not desirable to measure the reaction rate a t t = 0 because of induction time, initial stirring effects, temperature equilibration, etc. The intercept value could be reduced and the approximation of eqn. (8) could be improved by the more sensitive instrumentation, the more efficient stirring, and the automatic delivery as used by Malmtadt and hick^.^ In that case, A, is reduced to a minimum value consistent with induction time, etc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001652_j.jsv.2008.12.018-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001652_j.jsv.2008.12.018-Figure9-1.png", + "caption": "Fig. 9. Finite element model node locations.", + "texts": [ + " This system of equations was not sufficient to produce decent simulation results. This is due to the required small time step to get results from the system with the highest natural frequency of 44 kHz. Therefore development to simplify the system of equations to a level acceptable for simulation results was necessary. The development of the final simplified equations of motion included many steps starting with a linear modal transformation of separate front and rear driveshaft models using the node locations depicted in Fig. 9. Modal equations of motion were calculated using standard analytical modal analysis techniques to determine natural frequencies and mode shapes. These modal equations of motions were truncated to use only the equations with the natural frequencies in the range of study, modes with natural frequencies less than 500Hz. ARTICLE IN PRESS M. Browne, A. Palazzolo / Journal of Sound and Vibration 323 (2009) 334\u2013351342 The ITD was added to the front driveshaft using an additional degree of freedom coupled to the front driveshaft modal equations of motion at the node where it is located in the driveshaft", + " The jump is also apparent in the universal joint angles, reflecting the front shaft in the front universal joint, but a slightly different response in the rear shaft angle. The response of the rear universal joint angle is similar in shape as the response on the center bearing x-direction lining up with the rear driveshaft, and z-direction for the front driveshaft for the fourth-order response from the experimental test results. Comparisons of the experimental results in Fig. 12 to the predicted results in Fig. 9 shows some interesting features. Two peaks are present in both results, but the experimental measurement has the peaks closer together. These results align with the small differences between the normal modes and test results of the identification of the natural frequencies, see Table 2. Figs. 11b and 13 show a similar behavior with two peaks, and greater separation in the simulation results due to the results of the normal modes analysis. The simulation results are displacements while the measured results are acceleration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure16-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure16-1.png", + "caption": "Fig. 16 FEMs: (a) wave generator outer race (320 elements) and (b) flexspline (600 elements)", + "texts": [ + " The fraction of metal-to-metal contact decreases slightly as the input rotational speed and the environmental pressure increase. To clarify the lubricant film behaviour at the wave generator\u2013flexspline clearance, where the lubricating condition changed most drastically depending on the environment, a mixed lubrication analysis was conducted [4]. For considering the effect of elastic deformation, the outer race of wave generator and the flexspline were modelled by the finite-element method (FEM) as shown in Fig. 16. The inner race and balls of wave generator and the circular spline were treated as rigid bodies. The interaction between the races and balls of wave generator was assumed to be Hertzian contact. The load torque exerted to the flexspline outer surface due to teeth engagement with the circular spline was approximated by applying shear stress in proportion with the circumferential elastic deformation of the flexspline. Figure 17 summarizes the mixed lubrication analysis model and its coordinate system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002575_acc.2012.6315381-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002575_acc.2012.6315381-Figure3-1.png", + "caption": "Fig. 3. Illustration of one configuration for Case 2 to show \u2016\u03c8\u03b11\u2016 = \u2016\u03c8\u03b21\u2016: From (32), we have \u03c8\u03b21\u22a5b12, and \u03c8\u03b21\u22a5QTd b21 (top view). Since b12\u00d7b13 is normal to the plane P1 and the vector \u03c8\u03b21 lies on P1, we have \u03c8\u03b21\u22a5(b12\u00d7 b13). Similarly, for Pd2 , we also have \u03c8\u03b21\u22a5QTd (b21\u00d7 b23). In short, four vectors, namely b12, QTd b21, QTd (b21\u00d7 b23), b12\u00d7 b13, lies on the same plane perpendicular to \u03c8\u03b21 (top view). Therefore, \u2016\u03c8\u03b11\u2016 = \u2016QTd b21\u00d7 b12\u2016 = | sin \u03b812|, and \u2016\u03c8\u03b21\u2016 = 1 c \u2016(QTd (b21\u00d7 b23))\u00d7 (b12\u00d7 b13)\u2016 = | sin \u03b812|, i.e. \u2016\u03c8\u03b11\u2016 = \u2016\u03c8\u03b21\u2016.", + "texts": [], + "surrounding_texts": [ + "[1] D. Scharf, F. Hadeagh, and S. Ploen, \u201cA survey of spacecraft formation flying guidance and control (Part: II): control,\u201d in Proceeding of the American Control Conference, 2004, pp. 2976\u20132985. [2] M. Long and C. Hall, \u201cAttitude tracking control for spacecraft formation flying,\u201d in Proceedings of the Flight Mechanics Symposium, 1999. [3] A. Bondhus, K. Pettersen, and J. Gravdahl, \u201cLeader/follower synchronization of satellite attitude without angular velocity measurements,\u201d in Proceedings of IEEE Conference on Decision and Control, 2005. [4] H. Pan and V. Kapila, \u201cAdaptive nonlinear control for spacecraft formation flying with coupled translational and attitude dynamics,\u201d in Proceedings of IEEE Conference on Decision and Control, 2001. [5] J. Lawton and R. Beard, \u201cElementary attitude formation maneuver via leader-following and behaviour-based control,\u201d in Proceedings of the AIAA Guidance, Navigation and Control Conference, 2000. [6] T. Corazzini, A. Robertson, A. Adams, J. Hassibi, and J. How, \u201cGPS sensing for spacecraft formation flying,\u201d in Proceedings of the International Technical Meeting of the Satellite Division of the Institute of Navigation, 1997, pp. 735\u2013744. [7] P. Robert and R. Walkero, \u201cFixed wing UAV navigation and control through integrated GNSS and vision,\u201d in Proceedings of the AIAA Guidance, Navigation and Control Conference, 2005, AIAA 2005- 5867. [8] K. Gunnam, D. Hughes, J. Junkins, and N. Kehtarnavaz, \u201cA visionbased DSP embedded navigation sensor,\u201d IEEE Sensors Journal, vol. 2, no. 5, pp. 428\u2013442, 2002. [9] S. Kim, J. Crassidis, Y. Cheng, and A. Fosbury, \u201cKalman filtering for relative spacecraft attitude and position estimation,\u201d Journal of Guidance, Control, and Dynamics, vol. 30, no. 1, pp. 133\u2013143, 2007. [10] M. Andrle, J. Crassidis, R. Linares, Y. Cheng, and B. Hyun, \u201cDeterministic relative attitude determination of three-vehicle formations,\u201d Journal of Guidance, Control, and Dynamics, vol. 43, no. 4, pp. 1077\u2013 1088, 2009. [11] R. Linares, J. Crassidis, and Y. Cheng, \u201cConstrained relative attitude determination for two-vehicle formations,\u201d Journal of Guidance, Control, and Dynamics, vol. 34, no. 2, pp. 543\u2013553, 2011. [12] N. Chaturvedi, A. Sanyal, and N. McClamroch, \u201cRigid-body attitude control,\u201d IEEE Control Systems Magazine, vol. 31, no. 3, pp. 30\u201351, 2011. [13] F. Bullo and A. Lewis, Geometric control of mechanical systems, ser. Texts in Applied Mathematics. New York: Springer-Verlag, 2005, vol. 49, modeling, analysis, and design for simple mechanical control systems. [14] F. Bullo, R. Murray, and A. Sarti, \u201cControl on the sphere and reduced attitude stabilization,\u201d California Institute of Technology, Tech. Rep., 1995." + ] + }, + { + "image_filename": "designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003705_j.mechmachtheory.2015.07.004-Figure4-1.png", + "caption": "Fig. 4. Force on the wrapped cable segment.", + "texts": [ + " the co The wrapped angle will be decreased because of the bending rigidity of steel cable, as \u03b8e \u00bc \u03b8\u2212\u03b81\u2212\u03b82 \u00f09\u00de where \u03b81 \u2248 1 2 \u03c91 \u00bc 1 2R ffiffiffiffiffi B T1 s \u03b82 \u2248 1 2 \u03c92 \u00bc 1 2R ffiffiffiffiffi B T2 s 8>>< >>: \u00f010\u00de The effect of bending stiffness in the no-contact region can be concluded that the load inclines and the wrapped angle decreases. Inclining of the load could balance the torque caused by bending rigidity. While, the decreasing of wrapped angle is caused by the bow-out of the cable on the drum. The analysis of the force on the non-contact region is indispensable to solve the force equilibrium differential equation in the contact region, which will be discussed later. Fig. 4 shows a nonflexible cablewith bending rigidity passing around the drive drum.Amicro-section in contact region is considered, the force equilibrium in the tangential direction and the normal direction and the torque equilibrium around the center O is calculated in Eq. (11), Eq. (12), and Eq. (13), respectively. The inertia force of the fiber is neglected in the above force equilibrium equations. X F\u03c4 \u00bc T \u00fe dT\u00f0 \u00de cos d\u03c6 2 \u2212T cos d\u03c6 2 \u00fe Q \u00fe dQ\u00f0 \u00de sin d\u03c6 2 \u00fe Q sin d\u03c6 2 \u2212dF\u03bc \u00bc 0 \u00f011\u00de X Fn \u00bc Q \u00fe dQ\u00f0 \u00de cosd\u03c6 2 \u2212Q cos d\u03c6 2 \u2212 T \u00fe dT\u00f0 \u00de sin d\u03c6 2 \u2212T sin d\u03c6 2 \u00fe dN \u00bc 0 \u00f012\u00de X Mo \u00bc M \u00fe dM\u00f0 \u00de\u2212M\u2212 Q \u00fe dQ\u00f0 \u00deRd\u03c6 2 \u2212Q Rd\u03c6 2 \u00fe dF\u03bcr \u00bc 0 \u00f013\u00de where T, N, Q ,M and dF\u03bc denote the tangential force, the normal force, shear force, the bendingmoment and the friction force on the micro segment of the cable, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001491_09544062jmes1452-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001491_09544062jmes1452-Figure1-1.png", + "caption": "Fig. 1 Design of the rotor profile", + "texts": [ + " So, in the article, the rotor profile is formed by only two curves, which are arc and cycloidal curves. Using these two curves one can obtain a rotor profile with simple design, easy manufacturing, and fine sealing. Furthermore, a simple geometric design of the claw rotor can be derived without coordinate transformation and the meshing equation. The design method is easier than other common design methods. Finally, the article uses different parameters to discuss its influence on the gas port and pump performance. As shown in Fig. 1, the claw-type rotor consists of multi-piece curves. The equation of arc t1t2 can be represented as Rt1t2 = r sin \u03c6 \u21c0 i \u2212 r cos \u03c6 \u21c0 j (1) The angle range is 0 \u03c6 \u03b11. Curve t2t3 is a cycloid and the equation is Rt2t3 = [x cos(\u03b12 + \u03b13) + y sin(\u03b12 + \u03b13)] \u21c0 i + [y cos(\u03b12 + \u03b13) \u2212 x sin(\u03b12 + \u03b13)] \u21c0 j (2) where x = 2r cos(\u03c6 \u2212 \u03b11) \u2212 r cos(2\u03c6 \u2212 2\u03b11) y = 2r sin(\u03c6 \u2212 \u03b11) \u2212 r sin(2\u03c6 \u2212 2\u03b11) (3) The angle range is \u03b11 \u03c6 < (\u03b11 + \u03b12 + \u03b13). Curve t3t4 is an arc of radius R whose equation is shown as Rt3t4 = R sin \u03c6 \u21c0 i \u2212 R cos \u03c6 \u21c0 j (4) The angle range is (\u03b11 + \u03b12) \u03c6 \u03c0/2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003294_jjap.51.06fl17-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003294_jjap.51.06fl17-Figure3-1.png", + "caption": "Fig. 3. (Color online) Test structure for the evaluation of laser ablation. (a) Cantilever with an anchor. The scale bar is 10 m. (b) Determination of excess cutting length. During laser ablation, the femtosecond pulsed laser beam is focused at the bottom of the anchor.", + "texts": [ + " To remove the anchors attached to a microrotor, we constructed an optical system using a femtosecond pulsed laser (Spectra Physics Mai-Tai, wavelength 750 nm; repetition rate 80MHz; pulse width 100 fs). The laser beam is focused on each anchor using an objective lens with a numerical aperture of 0.65. Using this optical system, we examined the experimental conditions of laser ablation. As a test structure for evaluation experiments on anchor removal, we fabricated cantilevers with an anchor by twophoton microfabrication. The cantilever has the same anchor as the microrotor shown in Fig. 2. After fabrication, the cantilevers were metalized by electroless copper plating [Fig. 3(a)]. In the preliminary experiments, we found that part of the rotor blade was also removed by laser ablation of the attached anchor. For this reason, we examined the excess cutting length of the blade during laser ablation with different laser powers as shown in Fig. 3(b). Figure 4(a) shows the laser power dependence of the excess cutting 06FL17-2 # 2012 The Japan Society of Applied Physics length. In the experiments, the exposure time for each laser ablation was 0.5 s. The laser beam was focused at the bottom of the anchor as shown in Fig. 3(b). As a result, it was found that the excess cutting length was proportional to the laser power in the range from 50 to 250mW. However, at a laser power of 100mW, part of the anchor remained attached to the cantilever [Fig. 4(b)]. On the other hand, at a laser power of 150mW, the anchor was completely removed from the cantilever [Fig. 4(c)]. As a result, we removed the anchors of movable microparts at a laser power of 150mW in the following experiments. Using the above exposure conditions (laser power: 150mW, exposure time: 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003340_3.3786-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003340_3.3786-Figure1-1.png", + "caption": "Fig. 1 Slip surface location.", + "texts": [ + " Such problems generally involve the determination of a dividing stream surface, which separates injected fluid from freestream fluid and across which the pressure is continuous, but possible discontinuities in tangential velocity, entropy, temperature, and density may occur. The determination of this surface generally is rather difficult, involving the simultaneous solution of the external flowfield, between the slip surface and the freestream, and the injectant flowfield, between the slip surface and the body surface. In this note, we discuss a simple, exact solution of the inviscid equations of motion which includes the effect of surface mass transfer. Here the slip surface location is determined solely from the solution of the injectant flowfield (see Fig. 1). Consider the supersonic flow over a semi-infinite, slender, pointed cone of semi vertex angle 60. Mass is injected continuously from the cone surface at a constant rate of mass flow per unit area. It is assumed that the temperature of the cone is constant and that gas is injected with a velocity that is normal to the surface of the cone. (This last assumption Received October 11, 1965; revision received July 11, 1966. Any views expressed in this paper are those of the authors. They should not be interpreted as reflecting the views of The RAND Corporation or the official opinion or policy of any of its governmental or private research sponsors", + " (3) subject to condition 4 where = cos0 - M)] - 2} (5a) (5b) A = 2Mo(l - Mo2)1/2{m[(l + Mo)/(l - Mo)] - 2/Mo} (5c) B = -2/xo(l - Mo2)1/2 (5d) Equation (3) also arises in the study of constant density conical shock layers and is discussed on p. 143 of Ref. 2. This solution possesses the property that there exists an angle 6* for which Vi = 0. If v* = 0 at 6 = 6*, then from Eq. (5), 1 + \u00bb*i\u2014\u2014\u2014i1 - M* 2M* ^ __ , = m 1 + Mo i \u2014\u2014\u2014\u20141 - Mo Mo where M* = cos0* and MO = cos0o. The roots of this equation have been determined, and the dividing stream surface angle 6* is plotted in Fig. 1 as a function of the angle 00. We point out that the outer cone angle 0* is independent of the blowing rate ft, and in particular depends only on the fact that ft is not identically zero. There is no difference between suction and blowing in so far as the value of 0* is concerned; only the senses of streamlines are changed when ft changes sign. Compressible Inner Flow Equation (1) has been integrated numerically for various cone angles and injection Mach numbers M0. The integration is begun at 6 = do and is terminated where v\\ <10~13. This essentially corresponds to a flow that is in a radial direction. Figure 1 shows the variation of the slipcone surface angle with inner cone angle in air at a particularly high subsonic injection Mach number MQ \u00ab 0.92. The results for 0* for 7 = 1.4 and 7 = 1.67 differ by less than 0.3% and differ by only a few percent from the results of the calculation for the incompressible case. As in the incompressible case, the pressure at the slipcone is always larger than the pressure at the inner cone; the flow is always decelerated. It is clear, from the equations and initial conditions, that the slipcone angle 6* is independent of the blowing rate for fixed injection Mach number M0 and a fixed cone angle 00", + " Then the complicated flow interaction problem on secondary injection may be hopefully solved through the relatively tangible solid body situation. It is the intention of this note to consider this correspondence analytically. Explicit relationships are established between the dimension that characterizes the equivalent body and the injection parameters. Experimental data given in the recent literature are used for comparison with the analytical results and found to be in good agreement. Analysis Figure 1 depicts a flow model used in this investigation. It is assumed that a secondary fluid is injected into the supersonic stream and forms an equivalent body of certain shape. The secondary fluid may be either a gas or a liquid at high vapor pressure, inert or chemically reactive with the primary flow. The size of the equivalent body is determined by the exchange of momentum and energy between the secondary and primary fluids. First, considering the case in which no energy exchange is involved, it is assumed that the drag caused by the secondary fluid to the primary flow is the same as that from a slender body subject to the same supersonic flow condition" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003551_iros.2011.6094417-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003551_iros.2011.6094417-Figure2-1.png", + "caption": "Fig. 2: The head of the ARMAR-III humanoid robot", + "texts": [ + " Tracking experiments are presented in Sect. IV. Section V discusses also on-going research. 978-1-61284-456-5/11/$26.00 \u00a92011 IEEE 3192 The humanoid robot head used for the validation of the proposed gaze control concept has been developed within the SFB 588 \u201cHumanoid Robots\u201d. It has been designed to be used both as part of the humanoid robot ARMAR-III and as a stand-alone robot head for studying various perception tasks in the context of object recognition and human-robot interaction [16] (see Fig. 2a). The head has seven dof and its full kinematic scheme is shown in Fig. 2b. The neck motion is realized by four joints, with lower pitch q1, roll q2, yaw q3, and upper pitch q4 angles. The motion of the two eyes is realized by three more joints, with a shared common tilt axis (q5) and two independent rotations around a vertical axis (q6 and q7). All seven joints are driven by DC motors. The vision system consists of two cameras per eye, one with wide-angle lens for peripheral vision and one with narrow-angle lens for foveal vision. The used Dragonfly R\u00a92 cameras capture color images at a frame rate of up to 60 Hz" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003378_3.3995-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003378_3.3995-Figure1-1.png", + "caption": "Fig. 1 A shell of revolution composed of k conical-frustum shell elements.", + "texts": [ + " A series of trigonometric functions is assumed for each_of the three displacements 17, V, W and the four rotations J/XJ $e, i/V', i/V- Integration of the energy expressions would lead to integrals of the form/(I/a;) smcx dx and (If/x) coscx dx, which would result in infinite series. Thus, for computational convenience, a change of meridional-position variable first used by Mushtari and Sachenkov13 is used: x = xie y (14) where y is the new variable. Then the meridional positions of the small and large ends of the cone, x = xi and x = xz, are transformed to y = 0 and y = ln(#2/#i), respectively. To treat shells of revolution in addition to conical shells, the shell is approximated by a series of conical-shell segments, as shown in Fig. 1. This approximation has been used previously for static stress analyses14-15 and for vibrational analyses.10 Henceforth, the first subscript on the x dimension refers to the segment number, and the second subscript refers to the small or large ends, 1 and 2, respectively. Total strain and kinetic energies are found by summing over k, the number of conical-shell segments. Boundary Conditions Since there are three displacements and four angles of rotation of the normal to the middle surface, specification of boundary conditions could assume many forms" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001202_1.3179150-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001202_1.3179150-Figure8-1.png", + "caption": "Fig. 8 Balanced 2DOF parallel mechanism by combination of a balanced double pendulum and an idler loop", + "texts": [ + " The other then is still ecessary to balance the mass and inertia of the link itself. Since the angular velocities of parallel links are equal for the oment balance of two parallel links, only one CRCM is necesary. This means that there are only two CRCMs necessary, which oth can be constructed compactly near the base, as shown in Fig. . This is a configuration described in Ref. 14 . The former RCMs become fixed countermasses. It is also possible to derive this parallel mechanism by combiation of an idler loop 15 and a CRCM-balanced double penduum, as shown in Fig. 8. Also in this case, only two CRCMs are ecessary, and by using the countermass of the idler loop as a RCM, they can be constructed near the base. This result is hown in Fig. 9 and it has only one fixed countermass instead of wo, as in the configuration of Fig. 7. With the equations of the angular momentum of the balanced ouble pendulum being known, the inertias of the CRCMs and the nertia equations of the mechanism can be calculated quickly by imply adding the equations of each individual double pendulum" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002028_j.ymssp.2012.05.013-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002028_j.ymssp.2012.05.013-Figure7-1.png", + "caption": "Fig. 7. Stress distribution on tooth model.", + "texts": [ + " Since the starter/generator pad was equipped with a wide range loading capability, gear \u2018\u2018C\u2019\u2019 was selected for defect creation and remaining useful life and estimation. The fault was introduced after having performed a finite element analysis (using commercially-available software) on the gear in order to estimate the root bending stresses and to accurately determine the location of the maximum stress where a typical crack could initiate. The analysis was performed at different linear loads and were located close to the tooth tip and tooth mid-section. Centrifugal forces have also been accounted for in the FEM analysis. Fig. 7 shows the Von Mises stress distribution [27] for the case of a load of 134,150 N applied on the tooth tip and a speed of 3900 rpm. All studied cases showed a maximum stress located in the same vicinity. Based on these calculations, initial defects having 0.5 mm, 1.5 mm, 2 mm and 2.5 mm depth and 0.1 mm (0.004 in) width were created, each defect on a different spare gear (Gear C), at the gear tooth root, using electrical discharge machining (EDM). To reach a failure stage with the smaller defect, testing time would be very significant" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000165_bf00935252-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000165_bf00935252-Figure1-1.png", + "caption": "Fig. 1. G a m e kinematics and CCS orientation. P = pursuer , E = evader,", + "texts": [ + " Thus, 1, game of kind, (19) Ar~ = { _l/ t :r ' game of degree, is always positive in view of (12). It can be shown that the radial and angular adjoints &, A<, i = 1, 2, are given in terms of r and 0 as follows: ,L = 2% cos 0, (20) Ae, l +A,2 = Ar\u00a2r sin 0; (21) if the initial line (the reference with respect to which 0, B~, i = 1, 2, are measured) is chosen along the final LOS, Eqs. (20)-(2 t) simplify the angular adjoint rates to X,, = (-1)~A~,M~ sin ill. (22) Because of (7), Eq. (22) integrates in the CCS of Fig. 1 (with the origin coinciding with P 's terminal position and x-axis along the terminal PE) to A,, = (-1)'A~y,. (23) Since r and 0 are implicitly known in terms of (x~, y~), i = t , 2 (see Fig. 1), we see that (xi, yi, fii), i = 1, 2, can replace r, 0, ~b~, i = 1, 2, in a new canonical description of the game. The angular adjoints are proportional to the vehicle ordinates, by Eq. (23); the radial adjoint is given by (20). It is shown below that t i+t PMi = ( - -*) AMJAre, Yi and M/determine optimal strategies, and hence (x/, yi, M,, fi~, pc4,) result in a canonical description of the game which decouples into two disparate, identical sets, one for each vehicle. Later analysis brings out the advantages of such a transformation", + " Thus, a map of the paths emanating from the origin for different values of the terminal velocity vector (M~r, flit) represents the behavior of the aircraft both as pursuer and evader. This map is termed the extremal trajectory map (ETM). Remark 2.1. The position and orientation of CCS change from partie to partie; they are fixed only when the termination is known. Thus, (xi, yi) can only be calculated backward in time from termination. The relation between the reduced space quantities r, 4'i, i = 1, 2, and the CCS quantities is evident from Fig. 1. While constructing a reduced space trajectory from a pair of extremals, it can be easily checked whether r increases in retro-time \"r = t~- t. f+l , if z > O, s t e p ( z ) = ~indeterminate, if z = O, LO, if z < O. Remark 2.2. A reduced space formulation is chosen first, as is usual in differential game formulations. The nature of the radial and angular adjoints led to the choice of the CCS in this research effort. The conventional real-space formulation (with coordinate axes fixed to the plane and partie invariant) yields a canonical system of order 16 with no readily discernible relationships for the adjoints for further simplification" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000183_tmag.2008.2002996-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000183_tmag.2008.2002996-Figure1-1.png", + "caption": "Fig. 1. Flux line models.", + "texts": [ + " A time-stepping simulation is then possible; at each time step, the position of the rotor and the values of the feeding currents are updated and the airgap reluctance network is recomputed. Finally, in order to verify the effectiveness of the approach, the electromotive force (EMF) waveforms from the proposed nonlinear reluctance network models are, for different saturation levels, compared with those resulting from time-stepping finite-element simulations. According to a previous work [6], we propose the flux line models shown in Fig. 1. They are made of line segments and arcs in order to make the flux lines always normal to the air-ferromagnetic interface [5]. The flux density is assumed to be uniform along a flux line. These simple models permit us to define an airgap length function . The flux density is also assumed to be radial near the stator inner surface. Thus, the airgap length function is separable into two independent parts and , respectively, linked to the stator and the rotor, measured relatively to a borderline located in the middle of the airgap" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001200_00405160903178591-Figure27-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001200_00405160903178591-Figure27-1.png", + "caption": "Figure 27. Twisted yarn method.", + "texts": [ + " Figure 26 shows a schematic representation of the apparatus developed by Buckle and Pollitt [151]. The calibration of the apparatus plays an important role in the measurement; therefore, calibrating charts based on the yarn type should be used. The methods given above are primarily used for a yarn-to-metal friction test. For yarnto-yarn, the methods should be slightly modified. The static Capstan method can be used for yarn-to-yarn friction test by covering the surface of the cylinder with the same yarn. Another method is the twisted strand method, as shown schematically in Figure 27. The American Society for Testing and Materials (ASTM) has two standard test methods for yarn-to-yarn and yarn-to-metal testings. Exact procedure and dimension of the apparatus have been given in the ASTM standards [152,153]. In the previous sections, several experimental techniques and theoretical models proposed for the measurement of frictional characteristics of fibres were given. As mentioned earlier, the friction test is more sensitive than most of the physical tests; therefore, it would be necessary to consider the factors affecting the frictional characteristics of fibres" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002510_1.3616719-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002510_1.3616719-Figure11-1.png", + "caption": "Fig. 11 Compressor discharge pattern of tangential velocity variat ion. Data show steep gradient of tangential velocity near inner w a l l .", + "texts": [], + "surrounding_texts": [ + "Fig. 8 Compressor discharge diffuser. Locating the compressor discharge labyrinth air seal close to flow annulus inner w a l l eliminates the hazard of parasitic recirculating flows between whee l space chamber and diffuser Inlet.\nWhen a flow system has two or more diffusers operating in parallel, it is desirable that the critical flow ratio be always less than any operating conditions. Sonic velocity in turbine stators and jet nozzles tends to keep the flow ratios relatively constant for steady-state operating conditions. However, the flow ratio tends to be lower for some engine transient operations. For example, the high-frequency vibrations were most severe during rapid throttle advance from low engine speed.\nA large amount of diffusion in the duct increases the critical flow ratio. For this reason, transition ducts with large area ratios are more susceptible to instability than transition ducts with relatively low area ratios.\nIf the radial struts have adequate \"solidity,\" swirl in the discharge from a rotating blade row is effectively converted into equivalent static pressure rise in annular diffuser. With substantial inlet swirl entering a duct having constant-flow area normal to engine center line, there is a static pressure rise and the need to meet stability criteria in the design process. Reducing inlet swirl reduces the hazard of oscillatory flows.\nDiffuser Inlet-Wheel Space Chamber Coupling The compressor discharge air seal is usually located close to flow inner annulus wall as shown in Fig. 8. In other cases the labyrinth seal is located at substantially smaller diameter than wheel rim. A typical example is shown in Fig. 9. Advantages of this arrangement are reduced labyrinth seal leakage, and additional design flexibility as regards balancing of axial gas loads on rotor. When the wheel space chamber is substantial, a check should be made on possible crossflows and parasitic cellular recirculating flows.\nThe wheel space chamber coupled to the annular duct forms a compound resonator similar to the Boys double resonator. This is a combination of an organ-pipe type of resonator communicating near the closed end with an aperture leading to a chamber. The combination of aperture and chamber is a Helmholtz resonator whose resonant frequency is denoted by fa.\nThe combination has advantages as a sensitive instrument to measure sound waves. According to Rayleigh, \"the sensitiveness is exalted to an extraordinary degree. This is affected partly by the adoption of double resonance. The external vibrations may be regarded as magnified first by the large resonator and then again by a small one so that the mirror (indicator located in aperture) is affected by powerful alternating currents of air.\" Frequencies of resonant oscillation of the Boys resonator have been thoroughly investigated by E. T. Paris.\nWith turbomachinery wheel space chambers, powerful currents of air may flow in and out of diffuser inlet. Efficient diff users\nFig. 9 Compressor discharge annular d i f fuser\u2014where the compressor discharge seal is at a substantially smaller diameter than inner w a l l of main f low annulus, check for possible crossflow and cellular recirculating flows between w h e e l space chamber and diffuser inlet\nRADIAL STRUTS ANNULAR CHAMBER BETWEEN BLADE EXIT AND STRUTS,\u00a9\u2014\u2014\nWHEEL SPACE CHAMBERS-\nINFLOW1 REGION n = 1 CIRCUMFERENTIAL WAVES n=3 CIRCUMFERENTIAL WAVES\nFig. 1 0 Wheel space chamber coupled to diffuser inlet. Where an ax ia l gap connects whee l space chamber and diffuser inlet, cellular recirculating flows m a y occur.\nare known to be very sensitive to flows into and out of inlet, especially when this parasitic flow is injected perpendicular to or forward facing into the main stream. McMahan has thoroughly investigated the quantitative relationships [2].\nThe outer resonator of the Boys double resonator is open at one end. In turbomachinery, the \"organ-pipe\" resonator component is often closed at both ends. This is due to sonic velocity in blade bucket or rotor vane passages at both ends. Formulas for resonant frequencies of both types are given in the Appendix.\nFrequencies of resonant oscillations of the Boys double resonator determine the cut-off frequency for propagation of waves in a circumferential direction in annular chamber. In particular, the wheel space annulus has the properties of a highpass filter when an annular slot exists. As regards waves traveling in circumferential direction, the curved duct does not transmit waves having a frequency less than a critical or cut-off value determined by the frequency of resonant oscillations at right angles to the circumferential direction of propagation. The basic\nJournal of Engineering for Power O C T O B E R 1 9(5 7 / 5 1 7\nDownloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jetpez/26663/ on 02/06/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "RADIAL TRAVERSE\nDATA\nEXTRAPOLAT REGION 1\n\u2014 .\nON 1\n\u2014-a \u00bb\u2014 \u2014 \u2022 \u2022\u2014-\nHUB SPEED 1\ni iool ,\u2014i L_A 300 I 500 700 900 1100\nTANGENTIAL VELOCITY Cu, F t /sec\nTerman explains that \"A wave guide acts as a high-pass filter, and has a cut-off frequency determined by the dimensions of the guide and the particular field configuration involved. All waves having a frequency higher than the cut-off frequency will be propagated down the wave guide with the resultant transmission of energy, while waves of frequency lower than the cut-off value are rapidly attenuated in the direction of the axis of the guide and do not transmit energy. Furthermore, it is found that, under conditions where a wave will propagate down the guide, i.e., above the cut-off frequency, the velocity of phase propagation in ordinary types of guides will always be greater than the velocity of light (sound), corresponding to a velocity of energy transmission correspondingly less than the velocity of light (sound) as discussed below.\n\"The phase velocity represents the velocity with which an observer would need to travel along the axis of the wave guide in order to maintain a constant phase position with respect to the field configuration. The phase is the only thing that travels\nat this speed. The energy of the wave travels at the group velocity, which for the conditions that exist in a wave guide is related to the phase velocity and the velocity of light (sound) according to the relation\ne2 phase velocity = \u2014:\u2014-\ngroup velocity\n\"The group velocity is less than the velocity of light (sound) in the same ratio that the phase velocity is greater than the velocity of light (sound). At cut-olf frequency, where the phase velocity is infinite, the group velocity becomes zero and energy ceases to propagate down the wave guide. The component waves then bounce back and forth across the wave guide at right angles to its axis and do not travel down the guide at all.\"\nThe effect of a short communicating slot on frequencies of resonant oscillation of the two coupled annular conduits is to increase frequencies. This result was conclusively demonstrated by Irons in his studies of the fingering of wind instruments. Guided by electric transmission principles of circuits with distributed constants, analysis in the Appendix shows that the square of the frequency of resonant oscillation of combined system is equal to the sum of the squares of the frequencies of two component systems:\n\u2022 Frequency of circumferential waves with communicating slot. \u2022 Frequency of the two conduits and communicating slot reso-\nnating at right angles to circumferential direction. Detailed analysis is given in the Appendix.\nIn their experimental study of flow phenomena of partially enclosed rotating disks, Marati and his associates demonstrated periodic flow in the gap at the disk rim. Instead of a continuous radial inflow along the stationary sidewalls and a steady radial outflow along the surface of the rotating disk, the flow at any point around the parameter alternated periodically between radial inflow and outflow. Test data showed that the frequency of the radial flow reversal increased linearly with the rotational speed and that the number of flow regions tends to increase with increasing rotational speed, but decreases with increasing gap size.\nWith substantial gaps at disk rim, cellular recirculating flow into and out of wheel space chambers tends to have a rising pressure characteristic (one in which the local internal wheel space pressure increases with an increase of radial inflow). A state of equilibrium tends to be unstable because any momentary change in the flow alters the pressure in a direction tending to increase the disturbance. This results from the fact that tangential component of velocity in flow annulus is nonuniform. That boundary layer which is very close to the blade or bucket inner wall has a tangential speed approaching that of the wheel speed. The tangential velocity decreases rapidly at radial distance above the inner annulus. The radial pressure gradient opposing radial inflow is directly proportional to the square of the tangential velocity of the inflow. Within the wheel space chamber, the flow pattern is essentially a confined free vortex. That is, the tangential velocity of air varies inversely with the radius. The static pressure drop from flow annulus to wheel space is proportional to the square of tangential velocity. As a result, the internal static pressure in the wheel space increases with an increase of flow into the wheel space.\nOscillatory flows in wheel space chambers are demonstrated by repeated fatigue failures of thin walls of conical shells forming the inner wall of wheel space chambers. Typical failures are shown in Figs. 13 and 14. For example, repeated fatigue failures of turbine stage 3 rear labyrinth seal stator strongly indicate resonance. This seal stator is supported by a conical shell, which forms the inner boundary of the wheel space chamber aft of turbine stage 3. Strain gages put on the thin wall of this conical shell revealed a prominent flexural vibration at a frequency of 640 cps, and otherwise a low level of vibration activity. These data are shown in Fig. 15. Laboratory investigation on this conical shell showed that the natural flexural frequency for n = 3 waves in the\n5 1 8 / O C T O B E R 1 9 6 7 Transactions of the A S M E\nDownloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jetpez/26663/ on 02/06/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "Fig. 13 Turbine aft whee l space chamber inner w a l l fatigue cracks. Th in-wal l cone forming inner w a l l of wheel space chamber, which also supports aft air seal stator, has fatigue cracks (photograph no. 2864B2)\n1 STRAIN GAGE 14\nSTRAIN GAGE 11\n1000 2000 FREQUENCY.eps\nFig. 15 Turbine aft whee l space chamber inner wa l l . Vibration of thin wal ls of conical shell as measured on engine test.\nFig. 16 Turbine aft whee l space chamber inner wa l l . Frequencies of resonant flexural vibration of thin wal ls of conical shell determined by laboratory test, plotted on Campbel l d iagram.\n5 0 0 0\n4 5 0 0\n4 0 0 0\n3500\n3 0 0 0\n$ 2 5 0 0 z iij Z3 O UJ \u00a3 2000\n1500\n1000\nWITH STIFFNER\nWITHOUT STIFFNER\n2 0 0 300 4 0 0 5 0 0 6 0 0 ENGINE SPEED, rps\ncircumference was very close to 640 cps. Natural flexural frequencies of the thin walls of this conical shell are shown in Fig. 16. The mode corresponding to n = 3 circumferential waves has a frequency close to 640 cps. A relatively wide annular gap, having a width which varies from 0.31 in. minimum to 0.55 in. maximum, coupled turbine stage 3 aft wheel space chamber to inlet of the transition duct. Powerful flows circulate into and out of wheel space chamber, as demonstrated by repeated fatigue failures of conical shells inside wheel space chamber. Longitudinal acoustic waves in turbine discharge transition duct have a calculated fundamental tone close to the measured 640 cps at representative gas temperatures. Phase differences in longitudinal oscillations in adjacent passages result in apparent spinning pressure waves adjacent to buckets.\nLongitudinal oscillations having phase differences may excite blade and bucket vibration. Results of calculation for n = 3 circumferential waves rotating in forward direction are shown in Fig. 17(a). Resonance is revealed by calculations shown in Tables 1 and 2.\nStrain-gage measurements of buckets showed high vibration near 3 per rev excitation. Engine test data are shown in\nFigs. 17(6) and 17(c). Annular passages adjacent to leading edge and trailing edges of struts are coupled by parallel longitudinal passages as shown in Fig. 18. Phase differences in longitudinal oscillation in adjacent passages create apparent spinning pressure waves in both annular passages.\nFrequencies of resonant oscillatory flow around struts are affected by the area of constrictions between strut ends and adjacent bucket or stator vanes. The effect of changing area of constriction between strut exit and stator vanes is shown in Fig. 19. Reducing area of this constriction substantially reduces frequencies of resonant oscillation of higher modes but has little effect on modes having longest wavelength in axial direction.\nExtending strut fairings to leading edge of stator vanes helps prevent oscillatory flow around strut ends, and helps to stabilize flow. On aircraft with twin-inlet induction system, it is preferred and accepted practice to extend the splitter trailing edge to the leading edge of struts in engine front frame. A pliable seal is provided to accommodate assembly tolerances. The objective is to seal off the communicating channel between twin inlets.\nIf the diffuser or pipe terminates in nozzle, the end constriction\nJournal of Engineering for Power O C T O B E R 1 9(5 7 / 5 1 9\nDownloaded From: http://gasturbinespower.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jetpez/26663/ on 02/06/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use" + ] + }, + { + "image_filename": "designv11_3_0002591_10402004.2011.626144-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002591_10402004.2011.626144-Figure1-1.png", + "caption": "Fig. 1\u2014Bearing geometry and surface profile.", + "texts": [ + " Following the procedure of Paranjpe (5), the Reynolds equation is derived for the Maxwell\u2019s fluid in the same form as that of Eq. [1], except that the variable fluid-film pressure is p\u0304\u2217 instead of p\u0304. Using Galerkin\u2019s technique in a finite element method, the modified average Reynolds Eq. [1] is solved with the following boundary conditions: 1. Nodes situated on the external boundary have zero pressure, p\u0304|\u03b2=\u00b11 = 0. 2. Nodes situated on the supply hole have supply pressure p\u0304 = p\u0304s = 1. 3. At the trailing edge of the positive region p\u0304 = \u2202p\u0304 / \u2202\u03b1 = 0. Equations of Motion Referring to Fig. 1, the nonlinear equations of motion for a journal are [ M\u0304J 0 0 M\u0304J ] { \u00af\u0308XJ \u00af\u0308ZJ } = { F\u0304x \u2212 F\u0304ex F\u0304z \u2212 F\u0304ez } [5] where F\u0304x, F\u0304z are the unsteady-state reaction components due to hydrodynamic pressure and asperity contact pressure and F\u0304ex, F\u0304ez are the external loads in the x and z directions, respectively. The right-hand side of Eq. [5] represents the out-of-balance force components. It may be noted that the unsteady-state fluid-film reaction components due to hydrodynamic pressure are the functions of journal center displacements X\u0304j , Z\u0304j and journal center velocities \u00af\u0307Xj , \u00af\u0307Zj " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002116_j.camwa.2011.10.026-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002116_j.camwa.2011.10.026-Figure6-1.png", + "caption": "Fig. 6. Experimental hardware of PID controller, H\u221e controller and H\u221e hybrid controller.", + "texts": [ + " 5, so that they form a hybrid controller. Using the bilinear transformation s = 2 T z\u22121 z+1 |T=0.001 to replace s in (13), we can get the discrete expression of the hybrid controller shown as DHybrid(z) = Dr\u2212Hinf (z)D[Glead(s)] = 5088 \u2212 13750z\u22121 + 7284z\u22122 + 10050z\u22123 \u2212 12370z\u22124 + 3705z\u22125 1 \u2212 1.353z\u22121 \u2212 0.379z\u22122 + 0.6408z\u22123 + 0.2721z\u22124 \u2212 0.181z\u22125 . (14) In order to validate the different results of the three controllers and apply robust control to the EMA system, we carry out an experimental research. Fig. 6 schematically illustrates the experimental systemused, whichmainly contains an industrial control computer with a multi-functions interface card for data acquisition(12bit-16CH A/D) and control output(12bit-2CH D/A), a robust synthesized digital controller which runs under the real-time operating system, the actuator controller which contains the velocity loop, the electricity current loop, pulse-width-modulated(PWM) amplifier [17] and driving circuit, the EMA, and a torque spanner used to generate a disturbance torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002692_j.simpat.2012.06.001-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002692_j.simpat.2012.06.001-Figure9-1.png", + "caption": "Fig. 9. Reference voltage vector calculation in case j~Vref newj > j~Vbr j.", + "texts": [ + " 8), the reference voltage vector shown belongs in the determined twolevel hexagon. As a result the switching sequence followed for the first example is: 200\u2013210\u2013310\u2013311\u2013310\u2013210\u2013200. If j~Vref newj > j~Vbrj, then the determined two-level hexagon has to be changed. A new vector named ~V1;x, is used in order to determine the new two-level hexagon according to Eq. (19). In fact ~V1;x are the extra central vectors needed to be taken into account due to the increasing complexity. As a result both Eqs. (12) and (19) hold true. ~Vref new;2 \u00bc ~Vref ~V1;x \u00bc j~V2lref jejh2 \u00f019\u00de In Fig. 9, another example is shown. After determining the two-level hexagon it is observed that j~Vref newj > j~Vbr j. As a result the origin estimated vector ~Vref new needs to be changed in order to change the estimated two-level hexagon. The vector ~V1;x is used to calculate ~Vref new;2 according to Eq. (19). As a result the switching sequence followed for the second example is: 210\u2013310\u2013320\u2013321\u2013320\u2013310\u2013210. Note that the switching times calculation in both examples is according to Eqs. (13)\u2013(15). In Table 4 the limits of each sector are given" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003299_lindi.2012.6319476-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003299_lindi.2012.6319476-Figure5-1.png", + "caption": "Fig. 5. Visualization of distances with obstacles in the front.", + "texts": [ + " If (FrontDistance is Medium) and (LeftDistance is Far) and (RightDistance is Far) then (Angle is Negative) \u2022 7. If (FrontDistance is Medium) and (LeftDistance is Medium) then (Angle is Positive) \u2022 8. If (FrontDistance is Medium) and (RightDistance is Medium) then (Angle is Negative) In addition to this, we defined two additional fuzzy inputs to determine the distance of the obstacles that are positioned diagonally from the robot. The reason for this is that the primary objective for the car is to move forward in a straight line, so the obstacles in the front direction (Figure 4(a), Figure 5 marked as IF ) should have a greater weight than the obstacles on the side (Figure 4(b), Figure 5 marked as lL and IR). Also, we decreased the number of output membership functions to five (Figure 4(c)). Even in this case, it sometimes happened that as a result of different firing rules, the robot navigated itself to an obstacle on its left side (due to the firing of rule 5). An example to this error is if there is a close obstacle in the front and on the left, because in this case the robot will turn left (Figure 6), so this method cannot be used in every cases, it is advised to further improve it, which is described in the following part", + " The effects of the distances can be visualized using 3D graphs, where the z-axis is the output (the turn angle of the car, in degrees), while the x and y-axes are the distances (in pixels) according to the groups above. To examine how the changes in the third input variable affect the output (z-axis), the individual groups were tested under different circumstances. We changed the third input variable\u2019s value between three or four constant values, and this is how the effects on the output could be examined. For the different tests, these constant distances were 40, 18 and 2 pixels (in the case of the left and right sided distances; marked as IL and IR on Figure 5); while for the distances in the front, the constant values were 50, 35, 18 and 2 pixels (marked as IF on Figure 5). By using this method, the values of the two other input variables, and the value of the output variable can be read from the axes of the coordinate system. The correlation between the obstacles on the right side and in the front (with an approaching obstacle on the left side) is shown on Figure 7 (distance on the left side is 40 pixels on Figure 7(a), 18 pixels on 7(b), 2 pixels on 7(c)). The correlation between the obstacles on the left side and in the front (with an approaching obstacle on the right side) is shown on Figure 8 (distance on the right side is 40 pixels on Figure 8(a), 18 pixels on 8(b), 2 pixels on 8(c))" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001444_physreve.79.011906-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001444_physreve.79.011906-Figure4-1.png", + "caption": "FIG. 4. Decomposition of normal forces on a curved segment into resultant forces and couples of the normal force projections in two mutually perpendicular directions. a Sample profile of point normal forces fn on a curved segment. b Projected lengths Lx and Ly of the curved segment in \u2212x and \u2212y directions. c Projections of the point normal forces in the \u2212x and \u2212y directions. fn X is the x projection of the point normal force and fn Y is the y projection of a point normal force. d Decomposition of the projected forces into resultant forces and couples. Fn X and Fr X are the resultant force and couple due to the point normal forces projected in the \u2212x direction. Fn Y and Fr Y are the resultant force and couple due to the point normal forces projected in the \u2212y direction.", + "texts": [ + " In Appendix A we show that our idealization of the Brownian forces as point normal and point parallel random forces, and the consequent idea that diffusion coefficients capture the distribution of their resultants, is physically consistent with classical mechanics theory. By a treatment similar to that above, it can be shown that for a curved rod, the variance of the resultant of the Brownian point normal forces projected in a particular direction, is governed by the corresponding normal diffusion coefficient for the projected length in that direction see Fig. 4 . If fn is a point normal Brownian force acting at contour length l, and is the angle at that point, then the net projection of the point normal Brownian forces in the horizontal direction is \u2212L/2 L/2 fn cos l dl \u2212L/2 L/2 fn cos l dl = \u2212Lx/2 Lx/2 fndlx \u2212Lx/2 Lx/2 fn dlx = fn fn Lx = 4 KT 2 Dn X , 19 where Dn X is the normal diffusion coefficient corresponding to the projected length Lx. The above relation can be proved similarly for the parallel Brownian forces. Also, the variance of the total moment due to projected normal forces can be shown governed by the rotational diffusion coefficient for the projected length. For a projection in the horizontal direction, the net moment due to normal Brownian forces is 011906-5 \u2212L/2 L/2 l fn cos2 l dl \u2212L/2 L/2 l fn cos2 l dl = \u2212Lx/2 Lx/2 fnlxdlx \u2212Lx/2 Lx/2 fn lx dlx = fn fn Lx 3 12 = 4 KT 2 Dr X 20 and Dr X is the rotational diffusion coefficient corresponding to the projected length Lx. C. Large-deflection Euler beam response to resultant forces and couples in mutually perpendicular directions Consider the curved rod segment in Fig. 4, with N discrete point forces fn i acting normal to it at contour lengths li. Let i be the segment angle at the points of action of the forces. According to large-strain Euler beam mechanics, the moment balance at l=0 due to the N discrete forces is EI d dl 0 = i=1 N fn i cos i 0 li cos l dl + i=1 N fn i sin i 0 li sin l dl , 21 where l is the segment angle as a function of contour length, and each integral captures the moment arm of the corresponding force component. Now, the above equation is difficult to solve, and has been solved only for simple loading conditions using elliptic integrals 27 and, iterative and numerical methods 28\u201333 . However, by resolving the normal forces over the segment l=0\u2013L into resultant forces and couples in two independently perpendicular directions Fig. 4 d , the moment balance for that segment can be accurately determined at l=0. In Fig. 4, if \u2212x and \u2212y are the axes along which the Brownian forces are resolved,2 EI d dl 0 = Fn XLX 2 + Fn Y LY 2 + Fr X + Fr Y . 22 In Eq. 22 , Fn X and Fr X are the resultant force and couple of the x projections of the point normal forces acting on the segment Eqs. 23a and 23b , and Fn Y and Fr Y correspond to the y projections: Fn X = i=1 N fn i cos i, 23a Fr X = i=1 N fn i cos i 0 li cos l dl \u2212 LX 2 . 23b LX and LY are the \u2212x and \u2212y projections of the segment length Eqs. 23c and 23d , and are determined by numerical integration using a polynomial fit of l see derivation for Eq", + " First, they can be directly obtained from Eq. 25 by differentiating with respect to x and y, setting x and y to zero, and adding in the remaining equal and opposite forces that exist at x ,y=0 due to the couples Fr X and Fr Y. In Euler beam mechanics, \u2212EId2 l /dl2 physically represents an internal shear force that acts perpendicular to the beam at the point l. Secondly, Eqs. 26a and 26b can be shown to be true for an arbitrary distribution of point normal forces fn i , as in the case of the curved rod depicted by Fig. 4 and Eq. 21 . Consider stepping into the rod horizontally by an infinitesimal amount dx. At the new position, each vertical component of the point normal forces, fn i cos i, has decreased its moment arm by dx. Noting that the horizontal force components do not contribute to the moment about the horizontal lever arm, d dx EI d dl 0 = EI d dl 0 + dx \u2212 EI d dl 0 dx = i=1 N fn i cos i 0 li cos l dl \u2212 dx \u2212 i=1 N fn i cos i 0 li cos l dl dx = \u2212 i=1 N fn i cos i dx dx = \u2212 Fn X. 28 Finally, the difference between the end angles of the segment can be determined by integrating d /dl", + " Figure 9 a shows the normalized histogram of the angular fluctuations of the free end of the simulated chromosome. The distribution closely matched the predicted Boltzmann distribution Eq. 41 for the corresponding variance Eq. 42 . Also, the mean-square transverse fluctuations along the filament length showed the cubic dependence on filament length predicted by Eq. 43 Fig. 9 b . The time profile of thermal fluctuations in Fig. 9 c , however, suggested a smaller relaxation time than that observed in the experiments Fig. 4a in 35 . A preliminary calculation of the relaxation time gave an estimate of 0.14 s for chromosome fluctuations in water 36 , which was on the order of that observed in the simulations. Since the experimental observations were performed within a colchicine-arrested cell, and assuming cytoskeletal viscosity to be at least 100 times greater than that of water 2,37 , we repeated the simulation of chromosome fluctuation in a \u201ccell-like\u201d medium with 100 times the viscosity of water. The results for these simulations are shown in Figs. 9 d \u20139 f . The new time profile of the angular fluctuations showed an increase in the bending relaxation times Fig. 9 f , and was similar to that observed in the experiments Fig. 4a of 35 . Importantly, however, the dis- tribution of angular and transverse fluctuations Figs. 9 d and 9 e did not change with the change in viscosity, as predicted by the dependencies in Eqs. 41 and 43 . These results validate the ability of the model to capture the interplay between Brownian, viscous, and bending forces in the small-fluctuation or semirigid limit. We validated our model in the semiflexible limit Lp /L 1 by comparing the simulated thermal fluctuations of a phalloidin-labeled actin filament against the experimental observations of Gittes et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001070_10402000903283284-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001070_10402000903283284-Figure2-1.png", + "caption": "Fig. 2\u2014Section view of a rolling element bearing. Each of the objects i in the bearing, the inner and outer raceways, and the rollers has its own lubricant layer. The layer-thickness distribution is assumed to be equal to the average layer-thickness distribution h\u221e (s, t).", + "texts": [ + " The shape of the surface is defined by the radius function r(s). The flow of the layer on the surface is described using a local orthogonal coordinate system (s, \u03b8, n), where n is the coordinate in the direction of the outward normal to the surface. The thickness of the layer at a given position (s, \u03b8) and time t is denoted by h(s, \u03b8, t). A subscript i will be used to indicate that each solid object in the bearing, the inner and outer raceway, and each of the rollers has its own layer described in its own local coordinate system (Fig. 2). The continuity equation, as presented in van Zoelen, et al. (18), for a fluid layer on an axisymmetrical object (roller or raceway) labeled i reads, \u2202hi \u2202t = \u2212 1 ri \u2202 \u2202s (ri Qs,i) \u2212 1 ri \u2202Q\u03b8,i \u2202\u03b8i , [1] where Qs,i = Qs,i(si, \u03b8i, t) and Q\u03b8,i = Q\u03b8,i(si, \u03b8i, t) are the rates of volume flow per unit length and width of the layer in the si and \u03b8i directions, respectively. Next, a continuity relation will be derived that includes the lubricant layers on the raceways and on each of the rolling elements in one row in a rolling element bearing. In the bearing, multiple objects, numbered by index i, are in a configuration of rolling contact. It is assumed that there is no spinning motion. Due to the conforming geometry of the solid surfaces in si direction, a general coordinate s can be used to denote a position across the track on each object. This coordinate originates from the boundary of the track (Fig. 2). The average layer thick- ness in circumferential direction for position s is defined by h\u221e(s, t) = 1 lt nobjects\u2211 i=1 \u239b \u239dri 2\u03c0\u222b 0 hi d\u03b8i \u239e \u23a0, [2] where nobjects is the total number of objects in the configuration, and lt is the total length of the layers lt(s) = nobjects\u2211 i=1 2\u03c0ri(s). [3] D ow nl oa de d by [ M cM as te r U ni ve rs ity ] at 1 1: 04 2 5 N ov em be r 20 14 Equation [2] is differentiated with respect to time t, and Eq. [1] is substituted. Furthermore the term \u2202Q\u03b8,i/\u2202\u03b8i is omitted, as its integral with respect to the circumference is zero", + " The continuous exchange of fluid in the roller raceway contacts, together with other effects such as contact pressure and surface tension, will tend to distribute the lubricant evenly along the circumference of the rollers and the raceways. In addition, the flow speed is low compared to the rotational speed. Consequently, the change of the layer thickness in between two consecutive roller-raceway contacts can only be small. Therefore, in the long term after many overrollings, the layer-thickness distribution will be uniform in circumferential direction and equal for the raceways and the rollers. In Fig. 2 an illustration of such a layer-thickness distribution is shown. Direct evidence of the occurrence of equipartition for a single overrolling of a plate by a ring will be given in the Experimental Validation section. Indirect evidence for the occurrence of equipartition in starved EHL contacts is provided by van Zoelen, et al. (19). By assuming equipartition and a uniform layer thickness along the track, a model is obtained for the layer-thickness decay induced by contact pressure effects. This model was validated using a ball and disk apparatus showing good agreement also at high speeds and during many overrollings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000460_icpe.2007.4692551-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000460_icpe.2007.4692551-Figure1-1.png", + "caption": "Fig. 1. (a) matched Torque-Speed characteristics of two motor (b),(c) mismatched Torque-Speed characteristics of two motor", + "texts": [ + " This is shown in Fig. l(a). In this case, the torque of motor 1 is more than that of motor 2 and its speed is less. The difference between the speeds increases as the speed reference decreases. It is due to the fact that the difference between the rotor fluxes increases as the speed reference decreases. Now, if the parameters of the two motors have some tolerances and their characteristics are mismatched, and the same as before, the torque of motor 1 is higher, then the problem may get better or worse. In Fig. 1(b), the two motors have the same speed, however their torque is different. But in Fig. 1(c) the speed difference between the motors is more. In multi-machine single-converter systems, usually all the motors are the same and have identical characteristics. With induction motors connected in parallel, the load distribution is influenced only by the correct selection of the torque-speed mechanical characteristic. For squirrel-cage induction motors, no economical method for the adjustment of mechanical characteristic of the ready-made motors exists. For a slip-ring induction motor, the mechanical characteristic can be adjusted by including rotor resistors [8]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001049_j.snb.2010.03.069-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001049_j.snb.2010.03.069-Figure3-1.png", + "caption": "Fig. 3. Photographs of the patterned electrode substrate", + "texts": [ + " For creating AgCl on the Ag surface, a current of 50 nA was applied to the Ag electrode for 3 h [30,31]. The PDMS reservoir was made by molding, and an oxygen- and the fabricated reservoir-type oxygen sensor. 2 ctuat p c r t p 4 4 b o t e t p r r m A f ( F ( c w c u 66 J. Park et al. / Sensors and A ermeable membrane was either purchased or made by spin oating. The membrane was bonded onto the bottom of the PDMS eservoir opening. The patterned electrode substrate was bonded o the PDMS reservoir by using a silicone adhesive. Fig. 3 shows the hotographs of the fabricated reservoir-type oxygen sensor. . Results and discussions .1. Oxygen transport rate through the membrane In the oxygen sensor, the most important property of the memrane is its oxygen permeability because it is closely related to the xygen transport rate and the oxygen concentration gradient. In he oxygen transport rate experiment, PP, FEP, and PDMS were mployed as the polymer membranes of the fabricated reservoirype oxygen sensor with the WE area of 25,000 m2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003220_00423114.2011.621542-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003220_00423114.2011.621542-Figure9-1.png", + "caption": "Figure 9. Rider model for numerical simulations, for the sake of clarity only the stiffness element is represented.", + "texts": [ + " The tyre forces are applied on the actual contact point, which changes with roll angle and steer angle [16]. Mechanical compliances (lateral and torsion) of the mainframe and front frame are taken into account. Since road tests showed that in weave conditions upper body lean (\u03b8r) is relevant and lateral displacement (yr) is negligible, the rider\u2019s body model used for simulating weave phenomena has 1 dof. The lower body is fixed to the mainframe of the motorcycle, the upper body can perform lean rotation \u03b8r with respect to the motorcycle, see Figure 9. The biomechanical parameters of the model are: lower body mass, centre of mass position and moment of inertia about the longitudinal axis (these quantities are incorporated into the mainframe); upper body mass mur, centre of mass position hur and moment of inertia about the longitudinal axis IGur; torsion stiffness k\u03b8 rand viscous damping c\u03b8 rabout upper body lean axis; for the sake of simplicity the damper is not represented in Figure 9. Mass properties of the rider were calculated from anthropometric data. In particular, the upper and lower body masses were calculated according to References [17,18]. To identify the stiffness and damping properties of the joint between upper and lower body, some measurements of the lateral response of the rider to roll excitation were carried out in the laboratory with the methods described in [19]. The frequency response function (FRF) between upper body lean acceleration with respect to the motorcycle \u03b8\u0308r and roll acceleration \u03d5\u0308 was measured" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure8-1.png", + "caption": "Fig. 8. Helical rotating motion in a pipe", + "texts": [ + " If the wire is aligned and the wire is contracted, the shape of the body becomes a circular arc, as shown in Fig. 6. However, if the wire is fixed at an angle to the axis of the body and the wire is contracted, the shape of the body becomes helical, as shown in Fig. 7. Therefore, if four wires are attached to the body at an angle and the wires are contracted and elongated in turn, the body as a whole rotates on its axis. This is helical rotating motion. When this motion is achieved in a pipe, every part of the body rolls on the inner wall of the pipe, as shown in Fig. 8. Thus, helical rotating motion generates a propulsive force in a pipe. Ideally, no slippage occurs during locomotion because the direction of every roll axis is symmetric about the pipe axis, and all the body rolls at the same speed. In general, the more the wire is contracted, the larger the radius of the helical shape becomes. This section clarifies this relationship. We consider a cylindrical body with a radius of r, as shown in Fig. 9(a). A wire is attached along its body at an angle of \u03b1 relative to its flat face" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000294_1.2768069-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000294_1.2768069-Figure1-1.png", + "caption": "Fig. 1 Head-disk interface of the DTM recording system", + "texts": [ + " This is necessary because simulation for extremely small land and groove sizes are forbiddingly large numerical problems that cannot be handled on present-day computers. We compared the measured flying height losses due to DTM with the losses predicated by this simulation approach and with the losses predicated by an averaged flow model. We also used our simulation approach to evaluate the attitude of a slider flying over a DTM surface, the effect of groove parameters on flying profile, and the flying height losses due to manufacturing variation and changes at altitude. The interface between a head and a discrete-track disk is shown schematically in Fig. 1, where a number of grooves distributed on the disk are seen to be under or over the flying air bearing slider while the disk rotates. The spacing between the read/write head and the disk the distance from the read/write head surface to the top of the tracks on the disk must be maintained during HDD slider flying. Sliders flying over a discrete-track disk face a surface that consists of ridges and grooves, and the behavior of a slider flying over such a surface differs from that of a slider flying over a conventional smooth disk" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003140_1.4005032-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003140_1.4005032-Figure1-1.png", + "caption": "Fig. 1 Interaction between two gears in the process of normal mesh status", + "texts": [ + " In order to gain a better understanding of the interaction between oil film force and gear mesh force, lots of work needs to be performed. The purpose of this paper is to extend the previous studies further. This paper will consider the nonlinear oil film force and nonlinear gear mesh force, investigate the nonlinear dynamic responses of the system in detail, and compare the effect of nonlinear oil film force on the nonlinear mesh force and the effect of linear oil film force on the nonlinear mesh force. 2.1 Nonlinear Mesh Force. The interaction between the driving gear and driven gear is shown in Fig. 1. In Fig. 1, each gear has five degrees of freedom. They are yi; xi; hi; hyi; hxi(i \u00bc p; g). Here, p Contributed by the Technical Committee on Vibration and Sound for publication in the JOURNAL OF VIBRATION AND ACOUSTICS. Manuscript received March 7, 2009; final manuscript received August 16, 2011; published online May 29, 2012. Assoc. Editor: Alan Palazzolo. Journal of Vibration and Acoustics AUGUST 2012, Vol. 134 / 041001-1Copyright VC 2012 by ASME Downloaded From: http://vibrationacoustics.asmedigitalcollection", + " The relative gear mesh displacement is defined as D\u00f0t\u00de \u00bc rphp rghg \u00fe \u00f0xp xg\u00de sin a\u00fe \u00f0yp yg\u00de cos a e\u00f0t\u00de (1) Here, D\u00f0t\u00deis the relative gear mesh displacement, rp and rg are the radius of the base circle of the driving gear and driven gear, respectively, and a is the meshing angle. e\u00f0t\u00de is the static transmission error. The clearance is assumed to be 2bn, the nonlinear clearance displacement function f \u00f0D\u00f0t\u00de\u00de [10] can be defined as f \u00f0D\u00f0t\u00de\u00de \u00bc D\u00f0t\u00de bn;D\u00f0t\u00de > bn 0; bn D\u00f0t\u00de bn D\u00f0t\u00de \u00fe bn;D\u00f0t\u00de < bn 8>< >: (2) When two gears mesh with each other, the interaction force between gear teeth is shown in Fig. 1, and the dynamic mesh force relates with the vibration displacements of the two gears. Usually, the dynamic mesh force Fm\u00f0t\u00de is expressed as elastic force and damping force. Fm\u00f0t\u00de \u00bc k\u00f0t\u00def \u00f0D\u00f0t\u00de\u00de \u00fe cm _D\u00f0t\u00de (3) Here, k\u00f0t\u00de is the time-varying mesh stiffness, and cm is the mesh damping coefficient. The mesh damping can be calculated by cm \u00bc 2n ffiffiffiffiffiffiffiffi mkm p , where m is the equivalent mass of gear pair, m \u00bc IgIp=\u00f0r2 gIp \u00fe r2 pIg\u00de, and n is the damping ratio, n 0:06. The time-varying mesh stiffness and static transmission error are internal excitation, the input and output torque are both external excitation, and they can all be expressed in Fourier series form as k\u00f0t\u00de \u00bc km \u00fe Xn j\u00bc1 \u00bdk2j 1 cos\u00f0 jxmt\u00de\u00fek2j sin\u00f0 jxmt\u00de (4) The time-varying mesh stiffness and static transmission error are shown in Fig", + " e\u00f0t\u00de \u00bc XN j\u00bc1 ej cos\u00f0 jxmt\u00fe /j\u00de (5) Tp \u00bc Tp0 \u00fe XN i\u00bc1 \u00bdTi cos\u00f0ixpt\u00de \u00fe Ti\u00fe1 sin\u00f0ixpt\u00de (6) Tg \u00bc Tg0 \u00fe XM j\u00bc1 Tj cos\u00f0 jxgt\u00de \u00fe Tj\u00fe1 sin\u00f0 jxgt\u00de (7) Here, km is the average mesh stiffness of the gear pair, k2j and k2j 1 are the harmonious components of time-varying mesh stiffness, and ej is the harmonious components of static transmission error; however, the static transmission error is not the main point of this paper, just the first harmonic item of ej is used in the calculation. xm \u00bc ziXi\u00f0i \u00bc p; g\u00de, xm is the mesh frequency, /j is the harmonious phase item, zi is the number of teeth, Xi is the speed of the rotor, Tp and Tg are the input torques and output torques, respectively and Tp0 and Tg0 are the average value of the torques. According to Fig. 1, when the meshing state is in normal mesh process, that is, no tooth separation and back side collision, the excitations that act on the two gears include dynamic mesh force, torque, and unbalance. Therefore, the excitations that act on the two gears are defined as Table 1 Value of mesh stiffness and static transmission error Average mesh stiffness km(N=m) 6 108 Harmonic item of mesh stiffness k1(N=m) 2 108 Harmonic item of mesh stiffness k2(N=m) 1 108 Harmonic item of mesh stiffness k3(N=m) 4 107 Harmonic item of mesh stiffness k4(N=m) 2 107 Harmonic item of mesh stiffness k5(N=m) 5 106 Harmonic item of mesh stiffness k6(N=m) 2 106 First harmonic item of static transmission 10 10 6 Error e1(m) 041001-2 / Vol" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002160_2011-01-1548-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002160_2011-01-1548-Figure2-1.png", + "caption": "Figure 2. Loaded tooth contact analysis model: (a) Gear pair geometry, and (b) contact cells on engaging tooth surface.", + "texts": [ + " The gear mesh model can be obtained by several ways, including pitch cone design theory, unloaded tooth contact analysis and loaded tooth contact analysis [15, 16, 17]. The first two approaches use idealized tooth surface geometry only. The third one uses both detailed tooth surface geometry and the effect of material elasticity. In this study, the mesh model is condensed from the results of a loaded tooth contact analysis performed using a formulation that combines a semi- SAE Int. J. Passeng. Cars - Mech. Syst. | Volume 4 | Issue 21040 analytical theory with a 3-dimensional finite element (FE) approach [18]. The geometry of hypoid gear pair is shown in Figure 2(a), and Figure 2(b) shows the result of contact analysis qualitatively. For contact cell i in the mesh coordinate system Sm, its position vector is represented by ri(rix,riy,riz), contact force is represented by fi, and ni(nix,niy,niz) stands for its normal vector. A vector summation process is performed to obtain the total mesh force F given by (1) while the line-of-action N(nx,ny,nz) can be obtained from (2) The mesh point R(xm,ym,Zm) is computed by applying a method that is commonly used to compute the mass center of gravity of a group of rigid bodies" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure5-1.png", + "caption": "Fig. 5. (a) Five intersecting pairs form a closed double chain and (b) geometrical representation of its mobility conditions.", + "texts": [ + " The conditions for forming a closed loop are as follows. First, the end nodes, A and D should always meet when the mechanism is activated, and therefore, p1 \u00fe q2 \u00fe p2 \u00fe q3 \u00fe p3 \u00fe q4 \u00fe p4 \u00fe q5 \u00fe p5 \u00fe q1 \u00bc 0: \u00f03\u00de Secondly, nodes C and F should be connected, too. Thus, AC \u00bc DF: \u00f04\u00de The second condition is automatically satisfied because both AC and DF are equal to p1 q1. Now assume that the motion of the mechanism is determined by angle h( p 6 h 6 p). It is the angle between vector q1 and p1 and is positive clockwise as shown in Fig. 5a.2 The vector Eq. (3) can be replaced by two projections, along axes parallel with and perpendicular to vector q1. One of the projections is p1j j cos h\u00fe q2j j cos\u00f0 b1\u00de \u00fe p2j j cos\u00f0h a1\u00de \u00fe q3j j cos\u00f0 b1 b2\u00de \u00fe p3j j cos\u00f0h a1 a2\u00de \u00fe q4j j cos\u00f0 b1 b2 b3\u00de \u00fe p4j j cos\u00f0h a1 a2 a3\u00de \u00fe q5j j cos\u00f0 b1 b2 b3 b4\u00de \u00fe p5j j cos\u00f0h a1 a2 a3 a4\u00de \u00fe q1j j cos 0 \u00bc 0; \u00f05\u00de whereas the other is obtained by replacing cosine with sine. Rearranging Eq. (5) yields a cos h\u00fe b sin h\u00fe c \u00bc 0 \u00f06\u00de in which a \u00bc p1j j cos 0\u00fe p2j j cos a1 \u00fe p3j j cos\u00f0a1 \u00fe a2\u00de \u00fe p4j j cos\u00f0a1 \u00fe a2 \u00fe a3\u00de \u00fe p5j j cos\u00f0a1 \u00fe a2 \u00fe a3 \u00fe a4\u00de; \u00f07\u00de b \u00bc p1j j sin 0\u00fe p2j j sin a1 \u00fe p3j j sin\u00f0a1 \u00fe a2\u00de \u00fe p4j j sin\u00f0a1 \u00fe a2 \u00fe a3\u00de \u00fe p5j j sin\u00f0a1 \u00fe a2 \u00fe a3 \u00fe a4\u00de; \u00f08\u00de and c \u00bc q1j j cos 0\u00fe q2j j cos b1 \u00fe q3j j cos\u00f0b1 \u00fe b2\u00de \u00fe q4j j cos\u00f0b1 \u00fe b2 \u00fe b3\u00de \u00fe q5j j cos\u00f0b1 \u00fe b2 \u00fe b3 \u00fe b4\u00de: \u00f09\u00de To have a mobile linkage, the above equations must be satisfied for whatever h is", + " (3) along the direction perpendicular to vector q1 and a similar condition can be obtained simply by swapping cosine and sine in Eqs. (7)\u2013(9). A close inspection of first two equalities of Eqs. (10) and (1) reveals that geometrically they are equivalent to that vectors p1, p2,. . ., p5 form a closed loop. The first two equations from projection in the direction perpendicular to q1 lead to the same result. And the last equation, together with last equation from projection in the direction perpendicular to q1, as well as Eq. (2), indicates that vectors q1, q2,. . ., q5 form a closed loop, too. Both cases are shown in Fig. 5b. In other words, the projections to the axes parallel with and perpendicular to q1 shown in Fig. 5b give the exactly same expressions as Eqs. (1), (2)and(10). Moreover, if vectors p\u2019s and q\u2019s form closed loops, respectively, change of the expansion angle, h, does not alter them. Only the loops as a whole will rotate by the same amount as the variation of h, see Fig. 5b. Hence, we can conclude that the conditions that ensure the formation of a mobile closed loop double chain linkage consisting of five intersecting pairs are as follows: p1 \u00fe p2 \u00fe p3 \u00fe p4 \u00fe p5 \u00bc 0 \u00f011\u00de and q1 \u00fe q2 \u00fe q3 \u00fe q4 \u00fe q5 \u00bc 0: \u00f012\u00de The above derivation can be extended to double chain linkage consisting of n intersecting pairs by adding or reducing items in equations. The mobility condition for double chains with intersecting pairs is that the vector sum of edges of the pieces in every other pairs must be zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure9-1.png", + "caption": "Fig. 9 Inside of the test jig is observed by fiberscope: (a) Just after the start, (b) 441 rotational speed, (c) \u223c 33 200 rotational speed, and (d) the end of life", + "texts": [ + " The input torque began to rise from the vicinity beyond 33 000 output revolutions and increased 50 per cent of the initial value to 34 996 output revolutions. Then the test was stopped. Figure 8 shows the input torque at 90\u25e6 of the arm angle (maximum torque in the sinusoidal load cycle) during the life test. Some peaks of the input torque that are on the graph were caused by restarting the test after it has been stopped for the maintenance of equipment and were not used for judging the life test. Figure 9 shows the conditions of the input side tooth end and wave generator bearing in the test jig observed using a fibrescope. Wear conditions of each rubbing surface of the QM unit after the test are shown in Fig. 10. Proc. IMechE Vol. 222 Part J: J. Engineering Tribology JET415 \u00a9 IMechE 2008 at IOWA STATE UNIV on October 13, 2014pij.sagepub.comDownloaded from The remarkable wear is visible on each rubbing surface of the QM unit. The flexspline and wave generator were worn intensely on the input side and the wear reaches 60 \u03bcm at a deeper part" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001904_elan.201100442-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001904_elan.201100442-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of the experimental setup used to investigate ion-exchange at a resin bead using the CCD sodium ion image sensor.", + "texts": [ + " For the reverse experiment, 100 mL of 10 2 M sodium solution was placed on the sensor, and then 100 mL of 10 4 M sodium ion solution was quickly added. The CCD sodium ion image sensor monitored the solutions before and after each addition at 0.2 s intervals. Several beads of the sodium-type cation exchange resin were conditioned with 10 2 M sodium ion solution for least 12 h. At the same time, the CCD sensor was calibrated with three sodium ion solutions with concentrations of 10 1, 10 2 and 10 3 M. After these preparations were complete, a single bead of resin was placed on the CCD sensor as shown in Figure 1, and then covered with a thin paper sheet which prevented the bead from moving on the sensor as further solution was added. 90 mL of 10 2 M sodium ion solution was added slowly onto the CCD sensor. After the potential response was stable, 10 mL of 10 1 M of calcium or barium ion solution was quickly added. The CCD sensor monitored the ionexchange reaction before and after each addition. The images obtained every 0.2 s were analyzed. A multimicroscope (Nano Search Microscopena SFT3500, Shimadzu, Japan) was used to measure the size of the beads during the ion-exchange reaction in an experiment using the conditions described above" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002341_elan.201100265-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002341_elan.201100265-Figure1-1.png", + "caption": "Fig. 1. Steps involved in the preparation of the biosensor on GOx/Pt-NPs/graphene/BDD electrode.", + "texts": [ + " The diluted glucose solutions were prepared from the stock solution immediately before use. Sulfuric acid (H2SO4) and potassium hexachloroplatinate (IV) (K2PtCl6) were obtained from Aldrich Chemical (St. Louis, MO, USA). Nano diamond powder, with the average individual particle size of 3.2 nm, was purchased from JinGanfYuan New Material Development Co., LTD. Poly diallyldimethyl ammonium chloride (PDDA, MW= 400000\u2013500 000) and poly sodium 4-styrene sulfonate (PSS, MW =70 000) were obtained from Duksan Pure Chemicals (Gyeonggi-do, Korea). Figure 1 shows a schematic of the modified working electrode. First, the boron-doped nanocrystalline diamond film was deposited on the silicon substrate ( 1 in Figure 1). A thermally oxidized Si wafer was densely coated with the diamond nanocrystal seeds and then the doped diamond films were subsequently grown on the substrate [17]. In more detail, the diamond nanocrystal seeds were dispersed in a 10% (v/v) PSS aqueous solution by an attrition mill at 1000 rpm for 5 h, which resulted in cationic diamond nanocrystals coated with an anionic PSS chain. The thermally oxidized Si (100) wafer (oxide thickness: 1 mm) was coated with the thin cationic polymer, by dipping the substrate into a 10% (v/v) PDDA aqueous solution for 30 minutes and washing with DI water", + " Thin layers of nickel, with a thickness of less than 300 nm, were deposited on the SiO2/Si substrates using an electron-beam evaporator, and heated to 1000 8C inside a quartz tube under an argon atmosphere. After flowing the reaction gas mixtures (CH4:H2:Ar=50:65:600 at standard cubic centimeters per minute), the samples were rapidly cooled to room temperature (~25 8C) at the rate of ~10 8C s 1 using an argon stream. To transfer the synthesized graphene film, a Ni layer was etched using 1 M FeCl3 solution for several minutes, and the sample was rinsed with DI water. The floating graphene film was transferred onto a bare BDD substrate ( 2 in Figure 1). Electroanalysis 2011, 23, No. 10, 2408 \u2013 2414 2011 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim www.electroanalysis.wiley-vch.de 2409 The third step was the deposition of metal nanoparticles ( 3 in Figure 1). Pt-NPs were electrochemically deposited on the modified BDD electrode having a graphene sheet at constant potential of 0.6 V for 720 s in a 3-mL solution containing 20 mM K2PtCl6 and 0.5 M H2SO4. All the amperometric measurements were performed using an auto-lab microAUTOLABIII (Eco Chemie, The Netherlands) connected to a three-electrode cell. The three-electrode system consisted of a platinum (Pt) wire as a counter electrode, a SCE reference electrode, and the modified BDD thin film as a working electrode. The resultant electrode was denoted as a Pt-NPs/ graphene/BDD. GOx was then directly adsorbed onto the surface of the Pt particles by dipping them into a solution containing the enzymes shown in 4 of Figure 1. After the enzymatic layer was formed, the electrode was rinsed with a stream of a PBS to remove the residual monomers or the weakly linked enzyme molecules. Figure 2 shows the SEM images of the Pt-NPs electrodeposited at various potentials on the BDD electrode modified with graphene. The Pt nanostructures were electrochemically deposited in a solution of 20 mM K2PtCl6 and 0.5 M H2SO4 for 720 s. When a greater negative voltage was applied until it reached 0.6 V (Figure 2a\u2013c), the PtNPs became broader and denser" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001400_00423110903126478-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001400_00423110903126478-Figure4-1.png", + "caption": "Figure 4. Planar oscillator with two-dimensional friction damping.", + "texts": [ + " It has been shown in [9] that the friction force of the element is described by the following differential equation: T\u0307 = \u23a7\u23aa\u23a8 \u23aa\u23a9 K1(\u03bd\u0307 \u2212 Z\u0307) if |T | < T0, \u2212[\u2212K1(\u03bd\u0307 \u2212 Z\u0307)]+ if T = +T0, [K1(\u03bd\u0307 \u2212 Z\u0307)]+ if T = \u2212T0, (7) where [u]+ = { u if u \u2265 0, 0 if u < 0 . This description needs to be supplemented by the determination of the friction force direction angle \u03b1 according to the above described principle. An advantage of the description by the differential equation is the possibility of the direct application of the model of friction to multi-body computer codes, oriented for vehicle system dynamics. 2.2. Numerical tests of the model In order to test the model numerically, a planar oscillator shown in Figure 4 is investigated in the case of the isotropic friction. The oscillator is composed of a point body of mass M suspended by springs to the stationary foundation. Between the body and the plate there is dry friction. The plate may move in the horizontal plane with prescribed displacements \u03be (t) and \u03b7(t), and the motion of the body may be excited by initial conditions. But first let us explain the determination of the friction force direction angle \u03b1 on the example of this oscillator. The slip direction angle \u03b2 is determined by the formulae sin \u03b2 = Y\u0307 \u2212 \u03b7\u0307 D , cos \u03b2 = X\u0307 \u2212 \u03be\u0307 D , where D = \u221a (Y\u0307 \u2212 \u03b7\u0307)2 + (X\u0307 \u2212 \u03be\u0307 )2", + " The main peak is at about 35 Hz. The main frequency of dither is about 45 times higher than the eigenfrequency of the pendulum. Such a case may be considered similar to the M-F dither in the sense of rail vehicle dynamics. Referring to a real-world situation, if the basic vertical eigenfrequency of the vehicle is say 2 Hz, then dither of 45 times higher frequency will have frequency 90 Hz, which is the M-F. The substitute mechanical model of the experimental set-up is the planar oscillator shown in Figure 4. The realised option on the set-up is Kx = Ky = k. The dither excited by the actuator is equivalent to the motion of the plate of the planar oscillator described by displacements \u03be (t) and \u03b7(t). A number of experiments exploring the influence of dither on damping and validating the mathematical model of friction was carried out. The results of some of them are described below. In those experiments, the specimen was made of Prolab 65 Polyurethan C.N. A simple method to excite one-dimensional friction on the experimental set-up is to displace the bob from the stationary position and let it go without any initial velocity", + " In other words, smoothing by the M-F dither is not a tribological, microscale phenomenon. This allows applying the model of friction to materials other than those investigated experimentally, particularly the materials used for friction dampers of freight wagons. Subsequently, the proposed model of friction is used to describe the simulation model of a freight wagon. 8.1. Introductory numerical test The introductory test consists in applying measured dither generated by the rolling contact of wheel and rail to the modified planar oscillator of Figure 4. The modified oscillator has springs connecting the point body with the plate. The inertial, stiffness and damping parameters of the modified oscillator have been chosen, so that they roughly correspond to those of the two-axle freight wagon. The aim of the test is to assess the influence of the real dither generated by the rolling contact on this simple system, where creep forces do not interfere with damping. The equations of motion of the modified oscillator have the form MY\u0308 + Ky(Y \u2212 \u03b7) = Ty, MX\u0308 + Kx(X \u2212 \u03be) = Tx" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002901_jfm.2013.53-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002901_jfm.2013.53-Figure3-1.png", + "caption": "FIGURE 3. (Colour online) In the limit when A 1, a small cross-section of a ring (a) inclined at an angle \u03b8 to the flow\u2013gradient plane can be divided into two parts: an infinitely long cylinder aligned parallel (b) to the flow direction and an infinitely long cylinder perpendicular (c) to the flow\u2013gradient plane.", + "texts": [ + " Let the radius of the ring non-dimensionalized with c be A, measured from the centre of the global coordinate system whose origin is at the centre of the ring to the centre of the cross-section as shown in figure 1(d). We define the aspect ratio of the ring as the ratio of the radius of the circular cross-section to the radius of the ring, i.e. \u03ba = 1/A. For A 1, one can locally approximate the ring as a straight infinite cylinder. Thus, a small part of the ring of thickness (A d\u03b8 ) can be modelled as an infinitely long cylinder held at an angle \u03b8 to the simple shear flow (figure 3a), where \u03b8 is the azimuthal angle measured from the vorticity direction in the flow\u2013vorticity plane. Using linear superposition of Stokes flow solutions, the flow over an infinitely long cylinder inclined at an angle (\u03b8) can be divided into two parts: (i) flow parallel to a long cylinder (figure 3b) with a cos \u03b8 component of the shear flow acting over it; and (ii) flow perpendicular to a long cylinder with a sin \u03b8 component of the flow acting (figure 3c). The problem is now reduced to the solution of the force acting on an infinite cylinder aligned with its axis in the flow direction as shown in figure 3(b) and an infinite cylinder aligned with its axis in the vorticity direction as shown in figure 3(c). The undisturbed flow fields for the two problems in the new coordinate systems are now given by: vx\u2217 = 0, vy = 0, vz\u2217 =\u2212y cos \u03b8, (3.3) vx\u2217 = y sin \u03b8, vy = 0, vz\u2217 = 0, (3.4) where all the length scales are non-dimensionalized using the radius (c) of the undeformed ring cross-section, and the velocity gradients are non-dimensionalized using the shear rate, \u03b3 , of the simple shear flow. In the following discussion stresses are non-dimensionalized with \u00b5\u03b3 . The above flows are simple shear flows with cos \u03b8 Aligning rigid particles 131 and sin \u03b8 as the new non-dimensionalized shear rates; z\u2217 is the local coordinate along the axis of the ring and r is the local radial coordinate", + " vx\u2217, vy = 0 with vz\u2217(r, \u03b2) the only non-zero component of the velocity. With flow parallel to the axis one can say that the pressure p = 0 since only variation in x\u2217 and y are allowed and the velocity is perpendicular to these directions. The momentum balance equation then reduces to \u22072vz\u2217 = 0. The flow applies a shear stress along the axis of the particle but no forces in the x\u2217\u2013y-plane. There is no net force per unit length in the z\u2217-direction but rather a torque per unit length acting in the vorticity direction due to a force dipole as illustrated in figure 3(b). 132 V. Singh, D. L. Koch and A. D. Stroock In case (ii) (figure 3c), the imposed velocity in the x\u2217-direction induces stresses and a pressure variation in the x\u2217\u2013y-plane and a net force per unit length in the y-direction. If we further decompose the shear in the x\u2217\u2013y-plane into an extensional and a rotational component, we may note that the rotational fluid flow cannot induce a force. For a general three-dimensional particle, a fluid rotation far from the particle characterized by the vorticity \u03c9 can induce a force F= H \u00b7\u03c9, (3.5) where the resistance tensor H must be a pseudo-tensor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001795_s11012-010-9319-7-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001795_s11012-010-9319-7-Figure2-1.png", + "caption": "Fig. 2 Example of elasticity problem setup: (a) Geometry of the deformable bearing pad, finite element discretization, and fluid pressure distribution; sketch of a finite element. (b) Colorcoded contours of displacement magnitude on the interior surface of the bearing pad", + "texts": [ + " Quantitatively, (5) is obtained for the film thickness [2]: hhydro(\u03b8, z) = c + e0 cos(\u03b8) \ufe38 \ufe37\ufe37 \ufe38 Intact bearing + z[\u03c8y cos(\u03b8 + \u03c60) + \u03c8x sin(\u03b8 + \u03c60)] \ufe38 \ufe37\ufe37 \ufe38 Modification for misalignment + \u03b4h(\u03b8) \ufe38 \ufe37\ufe37 \ufe38 Modification for wear (5) From dimensional analysis of the EHD problem, it follows that the physics depends on the length-todiameter ratio, L/D, the Sommerfeld number, S, and the bearing deformation coefficient, \u03c80. In the present work, the bearing pad is considered elastic, and is housed inside a rigid bearing shell (see Fig. 1). The deformable bearing pad is modeled with two layers of linear 3-D wedge elements featuring 6 nodes per element, see Fig. 2. The bearing pad is assumed to be simply supported along the exterior surface, i.e., the local displacement vector is taken {d} = 0. The element faces in contact with the lubricant film are subjected to normal loads due to the hydrodynamic pressure, calculated by (1). The corresponding bearing pad elasticity problem is solved with the ABAQUS FEM software [16]. We note that the present model can be also applied to the problem of a rigid bearing pad featuring a thick deformable coating in the interior pad surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002477_1754337111414485-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002477_1754337111414485-Figure1-1.png", + "caption": "Fig. 1 Sole structure", + "texts": [ + " In general, the shoes used for health care running (mainly at low speed) attach importance to cushioning to attenuate the ground reaction impact load in the contact phase, and stability to prevent overwork of the joints in the feet. On the other hand, shoes used for athletic running (mainly at high speed) attach more importance to another property called operability, which makes it easy to keep running with low energy consumption for a long time, rather than focusing on cushioning and stability. This study considers the design optimization of a sport shoe sole structure with the aim of improving its operability. Figure 1 shows the sole structure considered in this study. This sole consists of midsole and outsole, which are made of ethylene vinyl acetate (EVA) foam and rubber, respectively. Proc. IMechE Vol. 225 Part P: J. Sports Engineering and Technology at University of Ulster Library on January 4, 2015pip.sagepub.comDownloaded from In this study, both the midsole and outsole were assumed to be linear materials. The midsole was divided into three domains: forefoot (Nos 1\u20139), midfoot (No. 10), and rearfoot (No. 11), as shown in Fig 1. The design variables in the present optimization study were the elastic modulus in each foot domain of the midsole, while the midsole shape, outsole shape, and outsole material were kept constant. Therefore, the design variables summed to 11. This definition seems valid because the hardness of EVA foam used for the midsole can be controlled by changing its expansion ratio. Furthermore, the forefoot domain was divided into 333 = 9 subdomains, to each of which an elastic modulus was assigned as a separate design variable. In addition, buffer domains were set between adjacent subdomains and involved the arithmetic means of the elastic moduli assigned to them. This can represent a smooth distribution of elastic modulus, which is similar to real shoe midsoles. The number of divided domains and subdomains (i.e. number of design variables) would be larger if the designer could finely optimize the midsole structure design; however, a midsole structure with a higher resolution than Fig. 1 is almost impossible to obtain in a real-world manufacturing process. The present optimization study involved quantitative evaluation of the requirements for shoes stated at the beginning of section 2. The operability, which is a property to be improved in this study, is supposed to depend on shoe weight and sole stiffness. Shoes should be lighter for higher operability as running with heavier shoes often leads to overconsumption of oxygen by the runner. In a series of constant-speed running tests, Nishida et al", + " It is speculated that such non-linearity comes from a complex interaction of mid + outsole materials and foot loading, although it is still hard to explain the mechanism of the interaction concretely due to lack of information and investigation. Proc. IMechE Vol. 225 Part P: J. Sports Engineering and Technology at University of Ulster Library on January 4, 2015pip.sagepub.comDownloaded from The sole weight can be promptly evaluated by summing volume 3 density for all elements. In the midsole, element density is determined by its elastic modulus, which is defined as a design variable (Fig. 1), based on the material properties profile of EVA foam shown in Fig. 3 (axis values were normalized between these maximum and minimum values). This profile is represented by spline curves that interpolate the data points given under four different hardness conditions (shown by black circles in Fig. 3). In the outsole, on the other hand, element density was constant as well as its elastic modulus. Figure 4 shows a flowchart of the design optimization employed in this study. The details of each step in the flowchart are described in what follows", + " In addition, this study considered two optimization cases based on equivalent design problem formulations (single-objective constrained problem and multi-objective non-constrained problem). A comparison of the results between the two optimization cases showed a slight difference in the resultant optimal solutions. The results indicated the effects of the choice of constraint handling techniques on convergence speed. The optimal midsoles found computationally in this study could be actually manufactured; it would involve adjusting the hardness of foam materials used in the 11 subdomains (Fig. 1) separately by changing the expansion ratio of EVA foam, and then joining these materials together by thermal bonding. In addition, the optimal midsoles can be validated by actual biomechanical testing; propulsion impulse is measured by running on a force plate that can sense a runner\u2019s ground reaction forces, while sole stiffness is measured by tracking video markers attached to the outside of the sole and analysing its three-dimensional deformation [4]. Thus, real-world manufacture and validation of numerically-optimal midsoles are necessary as a future work" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001946_j.engfracmech.2011.07.012-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001946_j.engfracmech.2011.07.012-Figure3-1.png", + "caption": "Fig. 3. The asperity point load model. The xz-plane is the symmetry plane.", + "texts": [ + " Thus, non-graded surface material data was used. The homogenity assumption was valid as long as the cracks remained close to the surface. The case hardening also resulted in a compressive biaxial residual surface stress, rR. However, after running-in of the gear, the residual surface stress was close to zero in the rolling direction [1], hence rR was taken as 0. The gear contact in Fig. 1a was modelled as a cylindrical Hertzian pressure distribution with maximum pressure p0l = 2.39 GPa [10] and contact half-width al, see Fig. 3. Lubrication separates the gear contact surfaces. According to elastohydrodynamic (EHD) lubrication [13], the thickness of the lubrication film will normally exceed 0.1 lm. It ensures normal load transfer, but removes the tangential one. Hence, frictionless full film lubrication was assumed for the cylindrical contact. The three-dimensional asperity contact was modelled with a spherical Hertzian pressure distribution with maximum pressure p0p, tangential traction q0p and contact radius ap. The asperity was assumed to break through the lubrication film", + " The zone with high tensile surface stress in front of the asperity, i.e. when x > 0, was identified as a potential RCF initiation region. An initial crack perpendicular to the surface would be subjected to rx. The RCF crack path was modelled in the symmetry plane (y = 0). The rolling contact with an asperity is three-dimensional, but by considering the symmetry plane through the asperity, a two-dimensional crack view was motivated and two-dimensional fatigue crack growth was investigated for the symmetry plane, see Fig. 3. This simplification is valid for spur gears and fully lubricated helical gears if the transverse friction component for the asperity is neglected. Note that by transforming the FE contact pressures to Hertzian pressures, closed-form expressions exist for the stresses at all positions in the assumed half-plane [15,16]. The stress state from the asperity point load was superimposed onto the twodimensional stress state from the cylindrical contact. The resulting stress field was used to model crack propagation with Matlab (R2009b)", + " At this point, the gear contact with an asperity on the pinion surface has been transformed into a moving cylindrical contact pressure, fixed spherical contact pressure and fixed spherical contact traction both with varying magnitudes, see Fig. 6. Alternatively, the pressures could be integrated into line and point forces. The cylindrical contact was represented by a concentrated normal line force Pl \u00bc pp0lal 2 : \u00f02\u00de Similarly, the asperity contact was characterized with a concentrated normal point force Pp \u00bc 2pp0pa2 p 3 ; \u00f03\u00de and a tangential point force Qp \u00bc 2pq0pa2 p 3 ; \u00f04\u00de which was directed opposite to the rolling direction, see Fig. 3, as slip was negative and thus q0p < 0. The closed-form expressions for the stress fields are available as elementary contributions [15,16]. For completeness, these are presented in Appendix A. Three combinations of elementary loads were investigated. Firstly, the load consisted only of the normal asperity load (Pp). Secondly, the tangential component on the asperity was included (Pp and Qp). Thirdly, both asperity loads were superimposed onto the normal cylindrical load (Pp, Qp and Pl). For all three load cases, both concentrated forces and distributed pressures were considered, see Table 3", + " The elementary solutions gave high accuracy in the solutions while at the same time the computation time remained reasonable. The predictions for initiation and growth of RCF cracks followed accepted fatigue theory. The fatigue damage process was at both ends explained by the asperity point load mechanism and linear elastic fracture mechanics. The authors greatfully acknowledge financial support from The Swedish Research Council. The stress field in the symmetry plane was the superposition of elementary solutions for a half-plane with the cartesian coordinate system in Fig. 3. Observe that the cylindrical load was positioned at xd and a translation of the coordinate system was required. In the xz half-plane, a biaxial residual surface stress rR would be superimposed to rx. A.1. Cylindrical contact A normal concentrated line load, Pl (Flamant problem) at x = xd generates the following stress field [15]: rx \u00bc 2Pl p \u00f0x xd\u00de2z \u00f0\u00f0x xd\u00de2 \u00fe z2\u00de2 ; \u00f0A:1a\u00de rz \u00bc 2Pl p z3 \u00f0\u00f0x xd\u00de2 \u00fe z2\u00de2 ; \u00f0A:1b\u00de sxz \u00bc 2Pl p \u00f0x xd\u00dez2 \u00f0\u00f0x xd\u00de2 \u00fe z2\u00de2 : \u00f0A:1c\u00de McEwen (1949) [15] expresses the stress field below the Hertzian pressure in terms of m and n, m2 \u00bc 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi f 2 \u00fe 4\u00f0x xd\u00de2z2 q \u00fe f \u00f0A:2a\u00de and n2 \u00bc 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi f 2 \u00fe 4\u00f0x xd\u00de2z2 q f ; \u00f0A:2b\u00de where f is f \u00bc a2 l \u00f0x xd\u00de2 \u00fe z2: \u00f0A:2c\u00de The signs of m and n are the same as the signs of z and x xd respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001114_0951192x.2010.528033-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001114_0951192x.2010.528033-Figure1-1.png", + "caption": "Figure 1. D-H parameters.", + "texts": [ + " Thus, through sequential transformations, the robot TCP coordinate system can be expressed in terms of the robot based coordinate system as function of the D-H parameters. The robot joints are numbered from 1 to n starting from the base side based on the condition that the ith joint precedes the ith link. The coordinate system Ri\u00fe1 \u00bc (oi\u00fe1, xi\u00fe1, yi\u00fe1, zi\u00fe1) is attached to the end of link i. The position and orientation of Ri\u00fe1 with respect to Ri \u00bc (oi, xi, yi, zi) attached to the end of link i71 is completely defined by four parameters (yi, ri, ai, ai) known as D-H parameters (Figure 1). The zi-axis is selected to be the axis of joint i, which will have two common normals connected to it: xi between zi71 and D ow nl oa de d by [ FU B er lin ] at 0 5: 20 1 4 M ay 2 01 5 zi, and xi\u00fe1 between zi and zi\u00fe1. The distance measured along the joint axis zi between the common normals xi and xi\u00fe1 is called ri. The angle from xi to xi\u00fe1 measured about zi axis is called ai. The shortest distance between the joints\u2019 axes zi and zi\u00fe1 of joints i and i \u00fe 1, respectively, measured along the common normal xi\u00fe1 is called ai" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002973_s0373463313000556-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002973_s0373463313000556-Figure1-1.png", + "caption": "Figure 1. AUV Coordinate systems.", + "texts": [ + " A comparison between the novel hybrid control strategy and the conventional backstepping control method is presented in this section. Finally, the work is concluded in Section 6. 2.1. Kinematic Modelling of AUV. It is normal to classify an AUV\u2019s coordinate system into a global coordinate system and a local coordinate system. The spatial position and orientation state vector in the global coordinate system can be expressed as \u03b7. The spatial linear velocity and angular velocity state vector in the local coordinate system can be expressed as q. The three dimensional AUV coordinate systems are shown in Figure 1. We consider that all kinematics equality constraints are independent of time. The position state vector \u03b7 and the velocity state vector q have the following relations: \u03b7\u0307 = J(\u03b7)q (1) where J[ R6\u00d76 represents the transformation matrix from the local coordinate system to the global coordinate system. It has been discussed in many papers (Dongkyoung, 2011; Serdar et al., 2010). 2.2. Dynamics Modelling of AUV. The three dimensional six degree-of-freedom dynamics equations of AUV can be expressed as follows (Dongkyoung, 2011; Serdar et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001942_j.electacta.2011.01.071-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001942_j.electacta.2011.01.071-Figure1-1.png", + "caption": "Fig. 1. Scheme and pictorial representation of the", + "texts": [ + " The AFM images ere recorded with a multimode scanning probe microscope sysem operated in tapping mode using Being Nano-Instruments SPM-4000, Ben Yuan Ltd. (Beijing, China). Synchrotron XRD paterns were recorded with a large Debye\u2013Scherrer camera installed t beam line 01C2 ( = 1.033 A\u030a or 12 keV) of National Synchrotron adiation Research Center (NSRRC) in Taiwan. The flow injection nalysis (FIA) system consisted of a high pressure microprocesor pump drive (BAS, USA) and a Rehodyne 7125 sample injection alve (20 L loop) [3,5,10]. A flow cell (\u223c50 L of volume capacty), as shown in Fig. 1, was used for analytical measurements. It onsists of an annealed-AuBPE working electrode, a stainless tube ounter electrode (outlet), and an Ag/AgCl reference electrode. All he experiments were performed at room temperature (25 \u25e6C). .2. Electrode preparations The AuBPE three-electrode system is the same as reported in ur previous studies [6,10]. The AuBPE (1.25 mm diameter, 31 mm ength) with an average weight of 392.4 \u00b1 0.6 mg (n = 10) was fabicated by barrel plating nickel thin film onto a copper rod and ubsequently plated gold thin film" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002184_cp.2012.0275-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002184_cp.2012.0275-Figure1-1.png", + "caption": "Figure 1: Stator cross section with non-overlapping windings.", + "texts": [ + " By utilising the optimal 3rd order harmonic, the average torque of two 10-pole/12-slot machines can be increased by ~11%, while their torque ripple is only slightly increased by less than 1.5%, compared to machines employing magnet shaping techniques without the 3rd order harmonic. The results are finally validated by finite element method. In order to evaluate the contribution of additional third order harmonic of PM shape to the performance, the 12-slot 10-pole SPM machine with two different rotor structures is adopted. The machines with non-overlapping windings have the same airgap length and share the same stator, as shown in Figure 1, and its main specifications and parameters are given in Table 1. The NdFeB PMs are employed and its remanence and relative permeability are also shown in Table 1. As stated above, the magnet shape can be optimized to achieve sinusoidal airgap flux density. The inverse cosine airgap length and sinusoidal shape magnets can be easily determined by analytical method, which thus will be employed in this paper. The airgap length is varied with the angle from the middle of magnets and, thus, the magnet shape produces sinusoidal flux distribution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000632_13506501jet415-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000632_13506501jet415-Figure3-1.png", + "caption": "Fig. 3 Configuration of SHF-20-160-2A-GR-SP", + "texts": [], + "surrounding_texts": [ + "Table 1 shows the requirements of SWG for this development. Considering that the major application of SWG for space applications is a paddle-drive and an antennapointing mechanisms, the SHF-type SWG with a hollow-shaped shaft was selected as a product for the development (Figs 2 and 3). It is made of stainless steel to prevent rust (Table 2) and its size is 20 (catalogue model number), which is currently the most used for space applications. Its reduction ratio is 160:1, which is the largest ratio among standard ratios and was selected because it is generally operated at very low The teeth of circular spline and flexspline, between the inner part of flexspline and outer part of wave generator and between races and balls of wave generator bearing, is lubricated by a multiply alkylated cyclopentane (MAC) grease. The bearing ball separator of wave generator is made of cotton-based phenolic resin, impregnated with MAC oil in vacuum. The grease application part and the quantity are presented in Table 3." + ] + }, + { + "image_filename": "designv11_3_0001077_s12206-009-0344-1-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001077_s12206-009-0344-1-Figure4-1.png", + "caption": "Fig. 4. Finite element mesh of the bushing model.", + "texts": [ + " 2 this force is [ ( ) ( ) ] t i t t t t t j i K f b d \u03b4 \u03b4 \u03b4 \u03b4= \u2206 + + = \u2212 df f f & (20) The input for the revolute bushing joint is the same as the cylindrical joint, plus the stiffness defined in the tangential direction ( )tK \u03b4 and the corresponding damping coefficient b. The models of the bushing joints require that their stiffness is defined. Because the bushings are rubber type materials, their stiffness is nonlinear and characterized by functions that need to be identified. For the purpose, four test cases were conducted in the finite element (FE) program ABAQUS [8] with an FE model of the bushing element. The bushing is modeled with 1404 solid \u201chybrid\u201d elements denominated by C3D8H in the finite element code, as depicted in Fig. 4. A rigid discrete cylindrical plate is created and tied to the nodes of the outer surface. A fixed boundary condition is prescribed to nodes in the inner surface. By applying displacements or rotations in the external rigid plate and measuring the reaction force or moments, respectively, in the plate the nonlinear stiffness constitutive functions for the bushings are obtained. The FE program used has several constitutive laws for nonlinear elastic and finite deformation analysis. A polynomial form of the material law is used for the force calculations, as it provides good prediction the experimental deformation values [5]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001594_bf02575194-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001594_bf02575194-Figure1-1.png", + "caption": "Figure 1. Four diagrams explaining features of ciliated tubules. In (a) the cross-sectional shape of the tubule is illustrated, (b) the relation of the type of metaehronism to the direction of the effective beat and fluid movement, (c) the effective and recovery strokes of a cilium, and (d) the", + "texts": [ + " A brief derivation of the force exerted by a cilium is presented to help formulate the problem. We know that in Stokes flow the force exerted by a body on the fluid is directly proportional to its relative velocity to the fluid. We define the normal (3F~) and tangential (SFT) force elements in terms of the normal (VN) and tangential (VT) velocities as follows: ~ F N = C ~ V N ~s = CN(v .n ) 3s, (4) ~ F T = C T V T 3s = Cr(v.t ) 3s, where v is the local relative velocity and n and t are the normal and tangential directions of the cilium at some point P, respectively (see Figure 1). The coefficients C N and C r are defined as follows: 2Ir/~, CN = 7CT, Cr = log L/ro + k l (5) where 7 is a slowly varying number between 1.4 and 1.8, while L is the length ro a characteristic radius of the cilium, and kl is another slowly varying constant of 0(1 ) in magnitude. Thus, the force element may now be written as: ~ F = [C~(_v._n)_n + Cr(v._t)~ ~s, (6) which, on use of the properties of the idemfactor I and substitution for CN and C r and ( t = 9$/~s) yields: ~_F = l o g L/ro + kl ~ - b' - 1 ) N N v_" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001190_icelmach.2010.5608143-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001190_icelmach.2010.5608143-Figure1-1.png", + "caption": "Fig. 1 Conventional magnetic gear.", + "texts": [ + " Index Terms--3-D finite element method, cogging torque, magnetic gear, skew. I. INTRODUCTION AGNETIC gears have some advantages such as low mechanical loss and maintenance-free operation that are not observed in conventional mechanical gears. In addition, they operate as a torque limiter under overloaded condition, and magnetic gears are expected to be applied to a joint of the human robot. However, most of previous magnetic gears have a problem of insufficient transmission torque for practical use due to the narrow facing area between two rotors (Fig. 1). Recently, various types of new magnetic gears to solve the above problem were proposed [1]-[5], and a SPM-type magnetic gear employing magnetic harmonics comes to attract attention because of its high transmission torque density though it has a complex structure with multipole magnets as shown in Fig. 2. Some studies on a SPM-type magnetic gear have been carried out, but few papers concerning the cogging torque can be seen. This paper describes the transmission torque characteristics in a SPM-type magnetic gear" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001488_j.cnsns.2010.07.025-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001488_j.cnsns.2010.07.025-Figure1-1.png", + "caption": "Fig. 1. Mechanical model of stick balancing with parametric excitation.", + "texts": [ + " Stability is determined using the first-order semi-disretization method [24,25]. The outline of the paper is as follows. First the mechanical model is introduced in Section 2. Then, some special cases are considered in Section 3. Section 4 presents the semi-discretization algorithm to the stability analysis of the system. Section 5 deals with the analysis of the critical length for different forcing frequencies and amplitudes. The results are concluded in Section 6. The mechanical model under study is shown in Fig. 1. The stick is attached to the horizontal slide that moves periodically up and down together with the base according to the geometric constraint r cos(xt). The stick to be balanced is assumed to be homogeneous, its mass is m and its length is l. The mass m0 of the slide is assumed to be negligible relative to the mass of the stick. The general coordinates are the angular position u of the stick and the position x of the pivot point. A control force Q is applied on the slide in order to balance the stick" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000650_j.mechmachtheory.2008.02.013-Figure3-1.png", + "caption": "Fig. 3. Coordinate systems applied to rack-generated rotor profiles and sealing line.", + "texts": [ + " A set of data is listed in Tables 2 and 3 for the rack prototype to verify the validity and the continuity of the profile. 2 data of designed twin-screw rotors (units: mm, deg) Item Symbol Example Lobe number of male rotor z1 5 Lobe number of female rotor z2 6 Center distance between rotor axes Ac 90.0 Outer radius of male rotor ro1 63.6 Outer radius of female rotor ro2 50.4 Inner radius of male rotor rd1 39.6 Inner radius of female rotor rd2 26.4 Pitch helix angle of rotor b 46.0 Length of rotor screw part L 205.9 As shown in Fig. 3, the coordinate system Sc is attached and translated with the rack, and the coordinate systems S1 and S2 are attached and rotated with the male rotor and the female rotor, respectively. The lobe profile of the male rotor and the groove profile of the female rotor can be generated by the concave and convex sides of the rack respectively. The rack locus can be derived in the coordinate system S1 of the male rotor, which is rigidly attached to the male rotor, as given in the following: x1\u00f0h;/1\u00de \u00bc \u00f0xc sc\u00de cos /1 \u00f0yc rp1\u00de sin /1 y1\u00f0h;/1\u00de \u00bc \u00f0xc sc\u00de sin /1 \u00fe \u00f0yc rp1\u00de cos /1 \u00f08\u00de x2\u00f0h;/2\u00de \u00bc \u00f0xc rp2/2\u00de cos /2 \u00fe \u00f0rp2 \u00fe yc\u00de sin /2 y2\u00f0h;/2\u00de \u00bc \u00f0rp2 \u00fe yc\u00de cos /2 \u00f0xc rp2/2\u00de sin /2 \u00f09\u00de where /1 is the rotation angle of the male rotor, /2 = z1/1/z2 is the rotation angle of the female rotor, sc = rp1/1 is the rack displacement, and the position vector in the rack coordinate system Sc can be represented as rc = [xc,yc] = [xh + d,yh], further, d is the distance from the rack center line to the point A", + " (12) can be substituted into Eq. (9) to obtain the equation of the female rotor profile as the following equations: x2 \u00bc \u00f0nhx=nhy\u00dexc cos /2 \u00fe \u00f0rp2 \u00fe yc\u00de sin /2 y2 \u00bc \u00f0nhx=nhy\u00dexc sin /2 \u00fe \u00f0rp2 \u00fe yc\u00de cos /2 \u00f013\u00de where /2 \u00bc ny xc nxyc ny rp2 . The length of the instant contact line (sealing line) between the female and male rotors is an important design parameter for the twin-screw compressor. The three-dimensional instant contact line between rotors is easily obtained using the contact path in the transverse plane. As shown in Fig. 3, the equation of the instant contact line can be derived by transforming the equation of the male rotor profile in the fixed coordinate system Sf as follows: xf1\u00f0h;/1\u00de \u00bc x1 cos /1 \u00fe y1 sin /1 yf1\u00f0h;/1\u00de \u00bc rp1 \u00fe y1 cos /1 x1 sin /1 \u00f014\u00de where /1 can be substituted by the rotation angle solved by Eq. (12). Then the three-dimensional instant contact line can be derived from the transverse contact path as [xf1,yf1,zf1], wherein zf1 = (rp1 tank)/1 and k is the pitch lead angle of the rotor. In all, the rack profile consists of nine sections" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001927_s12541-012-0021-7-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001927_s12541-012-0021-7-Figure7-1.png", + "caption": "Fig. 7 Involute curve of worm wheel (Left: Litvin and Fuentes)14", + "texts": [], + "surrounding_texts": [ + "The friction coefficient which is obtained from the tribometer can be mathematically converted to the worm gear efficiency by equation,2 and then the worm gear efficiency according to the normal contact pressure is obtained. In order to predict the worm gear efficiency according to the output torque, the normal contact pressure must be converted to the output torque. Gear geometry and Hertz\u2019s Law were employed for the variable conversion." + ] + }, + { + "image_filename": "designv11_3_0001330_1.3179933-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001330_1.3179933-Figure1-1.png", + "caption": "FIGURE 1. (a) The contact geometry; (b) First-shell particle environment with angular exclusions.", + "texts": [ + " The angle of friction cos(Semin) Va=/3 (1) The exclusion angle <56>m,\u201e is n/3 in a monodisperse packing of spheres or disks. The mechanical equilibrium condition can be ex pressed as complementarity relations in terms of the mul ticontact force pdf\u2019s: C k \\ I k \\ ^ fa Pkkf = 0 and ^ ra x fa\\ a=1 J \\a=1 J The functions Pc and Pkkr f contain a rich amount of information about the the state of a granular system in terms of the fabric and force distributions condensed in the particle environments, and they evolve with the driving strain while keeping to satisfy the constraints" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000097_taes.2007.4383606-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000097_taes.2007.4383606-Figure1-1.png", + "caption": "Fig. 1. \u201cTop view\u201d of constant-velocity trajectory parallel to n\u0302. Locations of sensors 1 and 2 are ~p1 and ~p2, the centers for corresponding spheres of radius r\u03031 and r\u03032. Line of airplane trajectory is depicted by dashed line; ~n stands for n\u0302. Points of closest approach to sensors p\u03031 and p\u03032 are points of contact between trajectory and spheres. Distance between these points of closest approach (see (3)) equals s\u0303(t\u03032\u00a1 t\u03031) (t\u03031 < t\u03032). (In figure, this combination is denoted s(t2 \u00a1 t1).) Distance between sensors equals k~p2\u00a1~p1k (\u00b8 s\u0303(t\u03032\u00a1 t\u03031)). Here ~ri denotes displacement vector between points of closest approach and corresponding sensor location: ~ri = p\u0303i \u00a1~pi, yielding ranges of closest approach r\u0303i = k~rik, i = 1,2: radii of corresponding spheres.", + "texts": [ + " The intrinsic advantages of statistical solutions notwithstanding [3, 7], the present approach, the like time-difference-of-arrival (TDOA) method [7, Appendix A], is geometrical. Errors in r\u0303, s\u0303, t\u0303 and in the sensor coordinates are assumed to be negligible, and, accordingly, all sensors\u2019 estimates for s\u0303 are taken to be equal. Sensors are also assumed to share a common clock. The natural question of observability is \u201cWhat is the minimum number of sensors capable, in this context, of identifying the trajectory?\u201d Fig. 1 illustrates the capabilities of two sensors; corresponding spheres touching the trajectory are depicted. The points of closest approach, p\u03031 and p\u03032 are part and parcel of the trajectory. However, even if the latter were known, rotations about the axis connecting the locations of the sensors would plainly generate additional solutions, with the closest-approach points tracing circles on the spheres (Fig. 1). Thus, observability demands additional sensors. In Section IIIA two trajectories are obtained from the r\u0303s, s\u0303, t\u0303s, and coordinates of three noncollinear sensors. This ambivalence is consistent with the observation that the reflection of an acceptable trajectory in the plane of the sensors evidently constitutes another acceptable trajectory. In Section IIIB a unique trajectory is obtained from four sensors in general position. Formulae for these trajectories are derived in Section III. Section II introduces four parameters which specify lines in R3", + " Using spherical polar coordinates, n\u0302= I\u0302sin\u03bccos\u00c1+ J\u0302sin\u03bc sin\u00c1+ k\u0302cos\u03bc, 0\u00b7 \u03bc \u00b7 \u00bc, 0< \u00c1< 2\u00bc (6) where I\u0302, J\u0302, and k\u0302 denote unit vectors along the positive x-, y-, and z-axes, respectively, and where \u03bc and \u00c1 are obtainable, as follows, from the t\u0303is and s\u0303 and the sensor locations (reserving the r\u0303is for the second stage). Consider the line passing through the closest-approach points to two sensors. Condition 3b) ensures that the displacement vectors between 1164 IEEE TRANSACTIONS ON AEROSPACE AND ELECTRONIC SYSTEMS VOL. 43, NO. 3 JULY 2007 the ~pi and their corresponding points of closest approach are perpendicular to ~v (see, for example, the ~ris of Fig. 1). Rotation about the aforementioned line would, therefore, rotate the sensor locations in circles in parallel planes, and the corresponding axial component of the displacement vector between the ~pis has magnitude equal the distance between these planes. Therefore, from condition a), the following equalities hold for the n\u0302-components of \u00a1\u00a1\u00a1\u00a1! p1\u00a1p2 and of\u00a1\u00a1\u00a1\u00a1! p2\u00a1p3: s\u0303(t\u03031\u00a1 t\u03032) = (x1\u00a1 x2)sin\u03bccos\u00c1+(y1\u00a1 y2)sin\u03bc sin\u00c1 +(z1\u00a1 z2)cos\u03bc (7) and s\u0303(t\u03032\u00a1 t\u03033) = (x2\u00a1 x3)sin\u03bccos\u00c1+(y2\u00a1 y3)sin\u03bc sin\u00c1 +(z2\u00a1 z3)cos\u03bc (8) respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001927_s12541-012-0021-7-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001927_s12541-012-0021-7-Figure6-1.png", + "caption": "Fig. 6 Geometry of worm wheel surface (Left: Litvin and Fuentes)14", + "texts": [ + " In the case of a ZI worm, the worm is generated by a straightlined blade. The cutting edges of the blade are installed in the axial section of the worm. Henceforth, two straight lines of the worm, which are generated by the straight line of the blade, are considered to calculate the tooth flank equation of the worm.10 Consequently, the other principal curve of the worm at the contact point is cot 2 p p s x z r\u03b1 = \u2212 + (8) where, sp = length of two opposite tooth flanks on the pitch circle. Worm wheel (Fig. 6, 7): The equation of a worm wheel surface is determined by the hob size. The equations of the worm wheel surface are14 cos cos sin sin cos cos sin tan h h h x r u y r u z u r \u03b8 \u03bb \u03b8 \u03b8 \u03bb \u03b8 \u03bb \u03b8 \u03bb = + = \u2212 = \u2212 + (9) sin cos cos cos cos sin tan h h h x r u y r u z r \u03b8 \u03b8 \u03b8 \u03b8 \u03bb \u03b8 \u03b8 \u03bb \u03b8 \u03bb = \u2212 + = + = (10) cos cos sin sin cos cos 0 h h x r u y r u z \u03b8\u03b8 \u03b8\u03b8 \u03b8\u03b8 \u03b8 \u03bb \u03b8 \u03b8 \u03bb \u03b8 = \u2212 \u2212 = \u2212 + = (11) where, rh = radius of the hob. In fully conjugated worm gears, theoretical line contact occurs. In that case, the gear teeth should be processed by a hob whose generator surface is identical to the worm surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000178_s12008-007-0003-7-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000178_s12008-007-0003-7-Figure3-1.png", + "caption": "Fig. 3 FEM mode: a outer ring, b 3D FEM model of slewing bearing, c inner ring", + "texts": [ + " The benefit is high too if the influence of geomet- rical or physical parameters has to be investigated. Eventually, reducing the model with this technique affords the possibility to build local models which reveal much more efficient than models of half the structure. Leray [10] showed that the behavior of the joint of one ring is independent of the configuration of the other ring, via an experimental study. Sizing a slewing bearing assembly then reduces to one study for the outer ring (model of an angular sector of the outer ring and its mounting), and a second study for the inner ring (see Fig. 3a\u2013c). In that case, the key data to be evaluated is the load distribution Fe carried by the rollers. 3.2 Fe loading The load distribution Fe is highly dependent upon the stiffnesses of the mountings. The test bench includes cylindrical mounting tubes. This shape offers the advantage of filtering the induced load in the structure. This will be discussed further in Sect. 4. This design eventually makes the load distribution on the roller tracks closer to a sine distribution. In the case when the loading follows a sine distribution, the equivalent load (Fe) applied on the most loaded angular sectors (sectors of screw 1 and 19) is defined by Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001114_0951192x.2010.528033-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001114_0951192x.2010.528033-Figure8-1.png", + "caption": "Figure 8. (a) Unimate 9000, (b) kinematic diagram of Unimate 9000.", + "texts": [ + " Figure 17 shows a three-dimensional view of sections in the estimated WS taken at different levels after an obstacle is placed close to the robot base. Estimated WS of the PUMA robot with Table 1. Data of case studies industrial robots. Robot Id No. of joints Axis No. Joint type D-H parameters Joint motion range y r a a Min. Max. Unimate 2000 (Figure 7) 4 0 \u2013 0.0000 0.0000 0.0000 0.0000 \u2013 \u2013 1 R q1 1.2000 0.0000 1.5708 70.2443 3.3859 2 R q2 0.0000 0.0000 1.5708 1.1170 2.0944 3 P 0.0000 q3 0.0000 71.5708 0.9650 2.0260 4 R q4 0.0000 0.1000 0.0000 73.4121 0.2075 Unimate 9000 (Figure 8) 7 0 \u2013 1.5708 0.0000 0.0000 71.5708 \u2013 \u2013 1 P 0.0000 q1 0.0000 1.5708 0.0000 4.3000 2 R q2 1.1000 0.0000 1.5708 70.7854 1.5708 3 R q3 0.0000 1.0000 0.0000 0.3491 1.5708 4 R q4 0.0000 0.1000 1.5708 70.8727 0.8727 5 R q5 1.2500 0.0000 71.5708 76.2832 6.2832 6 R q6 0.0000 0.0000 1.5708 71.9199 1.9199 7 R q7 0.1000 0.0000 1.5708 74.7124 7.8540 Fanuc S-5 (Figure 9) 6 0 \u2013 0.0000 0.0000 0.0000 0.0000 \u2013 \u2013 1 R q1 0.8100 0.2000 1.5708 71.0472 4.1888 2 R q2 0.0000 0.6000 0.0000 0.4363 2.3562 3 R q3 0.0000 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001070_10402000903283284-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001070_10402000903283284-Figure5-1.png", + "caption": "Fig. 5\u2014Experimental setup. A steel plate with four grooves, each with different depths. The cylindrical raceway is manually rolled over a groove, which is filled with oil.", + "texts": [ + " This assumption is not crucial in the model, as a different partitioning can be used, although this will give different results. This occurs when the surface properties such as wettability and roughness are not totally different, assuming equipartition seems reasonable, certainly upon average in the long run. To provide some justification for the hypothesis, a simple experiment is carried out. The experimental setup consists of a cylinder and a steel plate with four grooves with different depths: approximately 5, 10, 20, D ow nl oa de d by [ M cM as te r U ni ve rs ity ] at 1 1: 04 2 5 N ov em be r 20 14 and 40 \u00b5m (Fig. 5). At the start of the experiment, the steel surfaces are thoroughly cleaned. One of the grooves is filled with ample oil, which is subsequently carefully leveled off, in order to get a smooth layer with at least the thickness of the depth of the groove. The clean ring is aligned and manually rolled over the groove, making a single revolution with a rolling speed of approximately 0.02 m/s. Subsequently, the thickness of the oil layer picked up on the raceway and of the layer that remained behind in the groove is measured using optical interferometry" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure13-1.png", + "caption": "Fig. 13. Photograph and structure of the prototype.", + "texts": [ + " (17) Therefore, the length l of the wire is given with respect to time t as l(t) = 2\u03c0 \u221a\u221a\u221a\u221a\u221a\u221a ( a\u2212 r cos(2\u03c0 t T ) )2 + b2 \u239b \u239c\u239d1 + r2 sin2(2\u03c0 t T ) a2 + b2 \u239e \u239f\u23a0. (18) The curve of this equation is plotted in Fig. 12. One curve is along (18), and the other three curves are obtained by shifting the phase by \u03c0/2, \u03c0 and 3\u03c0/2 respectively. If the lengths of the four wires are changed along these curves, the body can make a helical rotating motion. A prototype driven by motors and pulleys was developed. Figure 13 shows a photograph and the structure of the prototype, which consists of joints connected by springs. The body is 10 mm in diameter, and each section is 22 mm long. The prototype has ten sections, so the entire body length is about 220 mm. Four holes are made in the joint concentrically at points 2.5 mm from the body axis. Four wires pass along the entire body through these holes. The joint parts are rotated 45\u02da relative to neighboring ones, determining the angle of the wires. One end of each wire is fixed at the end of the body, and the other end is pulled, rather than contracting the wire, to bend the body. If one of the four wires is pulled, the body shape becomes helical, as shown in Fig. 13. Teflon tubes are inserted into the holes to decrease the friction between the wires and the joints, so the force of the wires is transmitted effectively. This structure seems different from that shown in the section II-C. However, this structure is chosen for ease of fabrication. The important point is that the wires are fixed on the body at an angle; therefore, this structure is essentially equivalent to the one shown in the section II-C. If the four wires are pulled and released in turn, the prototype achieves helical rotating motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003147_j.mechmachtheory.2013.06.011-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003147_j.mechmachtheory.2013.06.011-Figure6-1.png", + "caption": "Fig. 6. O", + "texts": [ + " Simple computations show that (1) \u03a31 0 has 4 real points, if a b 1 (2) \u03a32 0 has 4 real points, if a N 1 (3) \u03a33 0 has 4 real points, if a b b n the left: Modes V0 and V5 and their intersection \u03a30 1 in (\u03b81, \u03b84, \u03b82)-space. Modes V0 and V11 and their intersection \u03a30 0 in (\u03b81, \u03b84, \u03b82)-space. In both pictures ter values a = 0.5 and b = 1.4 have been used. (4) \u03a34 0 has 4 real points, if a N b (5) \u03a35 0 has 4 real points, if b N 1 (6) \u03a36 0 has 4 real points, if b b 1 (7) \u03a37 0 has 4 real points, if either b b a b 1 or 1 b a b b (8) \u03a38 0 has 4 real points, if either a b 1 b b or b b 1 b a (9) \u03a39 0 has 4 real points, if either a b b b 1 or 1 b b b a. From this the result follows, see Fig. 6. Note that this result has a curious and perhaps unexpected consequence. For parameter values max{1, b} b a the variety bV29 does not intersect any other components. This could be called an isolated mode. Hence to reach this mode from the other modes the mechanism should be first taken apart and then put together in a different way. Similarly bV39 is an isolated mode for max{a, b} b 1 and bV58 for max{1, a} b b. Let us now consider the case a = b \u2260 1. Note that cases a = 1 \u2260 b and b = 1 \u2260 a yield completely analogous results, so it is sufficient to analyze only one of these cases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000839_nme.1620121102-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000839_nme.1620121102-Figure2-1.png", + "caption": "Figure 2. Axisymmetric shell element; (a) Axisymmetric thick shell element; (b) Axisymmetric shell element represented by mid-surface; (c) Mass system at a node i.", + "texts": [ + " The approximate rotary inertia matrix can be written as: T Z ; = ( I i z i e y ; + 113;&i)&i + (Itz;&; + 1z3iOzi)8yi + ( I n ; & i + Iz3i@,i)e,i (17) I l l , 0 (18) Approximation 2 If we assume that the rotations Oxi, Byi and t9zi are of the same order of magnitude, then exi = eyi = ezi and we obtain The approximate rotary inertia matrix can be written as: Approximation 3 In approximation 1 the off diagonal terms are neglected, whereas approximation 2 does not specify how the off diagonal terms should be distributed on the diagonal. In this approximation, the off diagonal terms are distributed on the diagonal in proportion to the diagonal terms. Let and The approximate rotary inertia matrix can be written as: 1642 K. S. SURANA Consider an axisymmetric element as shown in Figure 2(a). Also consider that the element is subjected to axisymmetric loading. For such an element a typical node i has two degrees-offreedom, u,, ui. Following a procedure similar to the general shell element, we construct an element (Figure 2(b)) with nodes on the mid surface only. A typical node i of such an element has three degrees-of-freedom, ui, vi, ai; ui, ui being the global displacements in the x and y directions and ai being the rotation about the global z axis. LUMPED MASS MATRICES WITH NON-ZERO INERTIA 1643 The lumped masses mf and mk for a pair of nodes i, and ib for the element of Figure 2(a) correspond to a lumped mass m,, and an inertia I i i at node i for the element of Figure 2(b). Following the reasoning similar to that of general shell elements, the mass rn i i can be thought of as the resultant of a pair of masses mii/2 concentrated at the ends of a rigid and massless rod V3, (Figure 2(c)). The inertia Iii of such a system is given by 2 midi 4 I.. = - where ti is the shell thickness at node i. The lumped mass matrix [Mi] for node i can be written as: 0 where {Sly= [ui, u,, ail. It is obvious that a rotation about the z axis has no effect on this mass matrix and its diagonal nature remains unaltered. For axisymmetric shells with non-axisymmetric loading, the inertia term Ii i remains unchanged and therefore this formulation is applicable. The lumped mass formulation with non-zero rotary inertia presented here has been implemented in the \u2018NISA\u2019 computer program" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000779_bfb0109667-Figure2.1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000779_bfb0109667-Figure2.1-1.png", + "caption": "Fig. 2.1. Exact tracking control: if the vehicle is on the contour, oriented along the tangent, a control proportional to the curvature of the target contour can achieve perfect tracking. The control is, however, unfeasible because of uncertainty.", + "texts": [ + " Let us pretend for a moment tha t our vehicle was on the target contour and oriented along its tangent, so tha t ~1 --- ~2 = 0. Then it would be immediate to write an exact tracking control law. In fact, from (1.10) one sees tha t choosing w -- v~3 causes ~i(t) = 0 V t. Therefore, in the peculiar case in which the vehicle is already on the target contour and heading along its tangent , a control proport ional to its curvature, namely ~(t) = 027. v~-~x2 (0, t ) (2.4) is sufficient to mainta in the vehicle on the contourat all t imes (see figure 2.1). Needless to say, we cannot count on the vehicle ever being exact ly on the contour. However, causali ty and the non-holonomic constraint imposes tha t any feasible t ra jec tory must go through the current position of the vehicle, and it must have its tangent oriented along its x-axis. Since the target contour is known only locally through the measurements V ( x i , t ) , i = 1 . . . N , one could imagine an \"approximation\" to the target contour which, in addit ion to fitting the measurements ~'(xi , t) , also satisfies the two addit ional nonholonomic constraints (see figure 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000466_iros.2008.4650802-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000466_iros.2008.4650802-Figure5-1.png", + "caption": "Fig. 5. Loop production.", + "texts": [ + " This relationship can be obtained from a description of the knot and analysis of the knot. Thus, recognition of the rope by a visual sensor is not taken into account in this paper. Loop production is an operation that makes one intersection. Thus, the description of the intersection is given by: El \u2212 Er \u2212\u2192 El \u2212 C {+,\u2212} 1 \u2212 C {+,\u2212} 1 \u2212 Er (a) El \u2212 Er \u2212\u2192 El \u2212 \u0302 C {+,\u2212} 1 \u2212 \u0302 C {+,\u2212} 1 \u2212 Er (b) This skill is equivalent to Reidemeister move I in knot theory. The intersection sign depends on the direction of the loop. Loop production strategy The process of loop production is shown in Fig. 5. In this skill, a loop that serves as the starting point of the knot is produced on the rope. First, the rope is grasped by two fingers. A loop is produced by the two dimensional translational motion of the hand. There are two different types of loop production depending on the grasp type of the two fingers, as shown in Fig. 5. This difference is important for performing the actual knotting process by robot fingers. To achieve robust loop production, the hand motion should be controlled by real-time visual feedback [1]. Rope permutation is categorized into two types, depending on the grasp type. In the case shown in Fig. 6(a), the description of the intersection is the following:\u23a7\u23a8\u23a9El \u2212 C {+,\u2212} 1 \u2212 C {+,\u2212} 1 \u2212 Er \u2212\u2192 El \u2212 C {+,\u2212} 1 \u2212 C {+,\u2212} 2 \u2212 C {+,\u2212} 2 \u2212 C {+,\u2212} 1 \u2212 Er In the case shown in Fig. 6(b), the description of the intersection is the following:\u23a7\u23aa\u23aa\u23a8\u23aa\u23aa\u23a9 El \u2212 \u0302 C {+,\u2212} 1 \u2212 \u0302 C {+,\u2212} 1 \u2212 Er \u2212\u2192 El \u2212 \u0302 C {+,\u2212} 1 \u2212 C {+,\u2212} 2 \u2212 C {\u2212,+} 3 \u2212 \u0302 C {+,\u2212} 1 \u2212C {+,\u2212} 2 \u2212 C {\u2212,+} 3 \u2212 Er The difference between the two types of rope permutation depends on the grasp type used in the loop production" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002923_0954409712445115-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002923_0954409712445115-Figure2-1.png", + "caption": "Figure 2. Definition of lateral track shift force Y and vertical (Ql , Qr ; left and right, respectively) and lateral (Yl , Yr) wheel loads, viewed in the rolling direction of the wheelset.", + "texts": [ + " 13 where friction coefficients measured with a hand-pushed tribometer are presented. In the current simulations, is varied between 0.3 and 0.6. A \u2018nominal\u2019 friction coefficient of 0.4 is employed unless otherwise specified. Measures of deteriorating influence of passing vehicles Lateral track shift forces In the following, lateral track shift forces are evaluated as a 2metre moving average (Y2m) of the total lateral at NORTH CAROLINA STATE UNIV on April 22, 2015pif.sagepub.comDownloaded from force (Yl Yr) induced by the wheelset on the track (see Figure 2). This measure is commonly employed as a safety limit criterion (the Prud\u2019Homme limit, e.g. Refs. 14,15) and has also been connected to lateral track deformation [16]. High lateral track shift forces may further promote lateral buckling of the track (sun kinks; see Ref. 17). According to Ref. 15, a highest permissible track shift force can be taken as Y2m4K 10\u00fe 2Q0 3 \u00f02\u00de where K \u00bc 0:85 for freight wagons and Q0 \u00bdkN is the static vertical wheel force. Equation (2) yields maximum permissible track shift forces of 93:5 kN for the heavy haul wagon (Q0 \u00bc 150 kN) and 79:3 kN for the freight wagon (Q0 \u00bc 125 kN)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003312_j.jsv.2013.05.022-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003312_j.jsv.2013.05.022-Figure11-1.png", + "caption": "Fig. 11. Representative magnetic forces acting on rotor, where Fl\u03c2\u03c2\u03be\u019b imply the l\u03c2th tangential (\u03be\u00bc t) and radial (\u03be\u00bc r) force on the ith magnet (\u03c2\u00bcm, \u019b\u00bc i), or those on the jth tooth (\u03c2\u00bc s, \u019b\u00bc j), and (a), (c), (e), (f) forces add, torques cancel; (d), (g) forces cancel, torques add; (b) forces cancel, and torques cancel.", + "texts": [ + " Practically, there are time-spatial-varying forces within the air gap, but this work employs equivalently spatial-fixed and time-varying concentrated instead of distributed loads to model the magnet- and slot-frequency forces. Relatively, the Finite Element results have more practical meaning because the distributed effect of the loads are to-much-degree taken into account especially when very fine grid blocks are chosen. Note that the simulation results are well consistent with analytical ones, which in fact verifies the concentrated-load assumption and the superposition method. To illustrate more, Fig. 11 takes the 4-magnet/3-slot motor for example to address several possibilities of the force configurations, where Fl\u03c2\u03c2\u03be\u019b imply the l\u03c2th (l\u03c2 \u00bc 1;2;3 or 4) harmonic tangential (\u03be\u00bc t) and radial (\u03be\u00bc r) forces on the ith magnet (\u03c2\u00bcm, \u019b\u00bc i), or those on the jth tooth (\u03c2\u00bc s, \u019b\u00bc j). For the rotor-side forces, the phases at the first four example harmonics are {lm, Ns, Nm, 2\u03c0lmNs\u00f0i\u22121\u00de=Nm}\u00bc{(1, 3, 4, 0\u2218)i\u00bc1, (1, 3, 4, 270 \u2218)i\u00bc2, (1, 3, 4, 180 \u2218)i\u00bc3, (1, 3, 4, 90 \u2218)i\u00bc4, (2, 3, 4, 0\u2218)i\u00bc1, (2, 3, 4, 180 \u2218)i\u00bc2, (2, 3, 4, 0 \u2218)i\u00bc3, (2, 3, 4, 180 \u2218)i\u00bc4, (3, 3, 4, 0 \u2218)i\u00bc1, (3, 3, 4, 90 \u2218)i\u00bc2, (3, 3, 4, 180 \u2218)i\u00bc3, (3, 3, 4, 270 \u2218)i\u00bc4, (4, 3, 4, 0\u2218)i\u00bc1, (4, 3, 4, 0 \u2218)i\u00bc2, (4, 3, 4, 0 \u2218)i\u00bc3, (4, 3, 4, 0 \u2218)i\u00bc4}", + " Because of different configurations of the inner excitation phase, vibrations in PM motors can be classified into rotational modes where the net force is canceled but the torque is added, and thus motors only exhibit rotational rather than translational motion, translational modes where the net torque is canceled but the net force added, and thus motors has translational instead of rotational motion, and the balanced modes where the net force and torque are both canceled, and consequently motors have no rigid motion. The typical vibrations can be verified by the force phase configurations upon the magnets and teeth. In Fig. 11, (a) and (c) imply torque cancel and force add (ks \u00bc 3 and 1, respectively), and thus the translational modes are excited, whereas (b) means torque and force cancel, and thus the balanced modes excited (ks \u00bc 2), but (d) implies the torque add and force cancel, and thus the rotational modes are excited (ks \u00bc 0). For the stator, (e) and (f) imply torque cancel and force add, and thus the translational modes are excited (km \u00bc 1 and 2 respectively), whereas (g) implies torque add and force cancel, and thus the rotational modes excited (km \u00bc 0)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003347_j.proeng.2012.09.530-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003347_j.proeng.2012.09.530-Figure2-1.png", + "caption": "Fig. 2. Kinematics in the contact region", + "texts": [ + " The function )\u02c6(\u02c6 CW is given for the Mooney-Rivlin material )3\u02c6()3\u02c6()\u02c6(\u02c6 0110 \u2212+\u2212= III IICW \u03bc\u03bc (9) where 10\u03bc and 01\u03bc are the material constants, II\u0302 and III\u0302 are the first and second invariants of the unimodular right Cauchy-Green tensor defined as 332211 \u02c6\u02c6\u02c6\u02c6 CCCI I ++= , )\u02c6\u02c6\u02c6( 2 1\u02c6 CCII III \u22c5\u22c5\u2212= (10) In the reference coordinate system, the II. Piola-Kirchhoff stress tensor for the rubber is given by 1)\u02c6(\u02c6 2 \u2212+ \u2202 \u2202 = CJp C CW S (11) where p is the hydrostatic pressure. In the current configuration the Cauchy stress tensor can be calculated as TFSFJT 1\u2212= . (12) Let us consider a system which consists of two bodies. One is elastic signed by 1 , the other is a rigid body signed by 2 see in Fig. 2., where cA is the contact region. In this work normal contact is assumed, where cn is the normal unit vector of the contact surfaces through two point pairs 1Q and 2Q . The ng normal gap can be defined as huugg nnn +\u2212== 1)( , cAr \u2208 , (13) where h is the initial gap and 1 nu is the normal component of the displacement vector. There can be two cases, one is contact, when 00 \u2265= nn pg , (14) there is gap (no contact) between the two bodies if 00 =\u2265 nn pg , (15) where np is the contact pressure. Both cases 0=nn gp , cAr \u2208 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001981_tia.2012.2226551-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001981_tia.2012.2226551-Figure12-1.png", + "caption": "Fig. 12. PMAAC motor.", + "texts": [], + "surrounding_texts": [ + "The CSR front-end is a buck converter, as opposed to the VSR, which is a boost converter. That is, the dc output voltage of the CSR can only be lower than the peak input ac voltages, whereas the VSR output voltage can be higher. Since the VSI stage is also a buck converter, this means that the motor voltages that can produced by the CSR/VSI are lower than the input voltages. This characteristic is a disadvantage of the CSR/VSI when compared to the VSR/VSI, unless the motor requires a lower voltage\u2014in which case, it will actually be an advantage because the VSI voltage conversion ratio can be minimized and VSI conversion is accomplished with lower losses. The PMAAC motor, shown in Figs. 12 and 13, provides a power-dense solution when the application requires a high diameter-to-length ratio [6]. The PMAAC motor has an axial-flux and \u201cinside/out\u201d structure and, therefore, no stator back iron [7]. Commercially available PMAAC motors utilize printed circuit boards for the stator winding structure and are packaged in a \u201cstacked pancake\u201d form factor [6]. The PMAAC motor usually has a lower voltage rating because of tradeoffs between voltage isolation capability and cost. For a 440-V-fed system, a 250\u2013300-V motor rating is optimal in the 5\u201310-hp range." + ] + }, + { + "image_filename": "designv11_3_0002326_robio.2010.5723563-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002326_robio.2010.5723563-Figure14-1.png", + "caption": "Fig. 14. Task space model for the experiment.", + "texts": [ + " These procedure is repeated until the most appropriate candidate grasp is obtained [9], [11], and the grasping position and orientation of the candidate is inputted in a motion planner of the robot arm to achieve the grasping task. We applied our grasp planning method to a mobile manipulator system shown in Fig. 13, and required the system to grasp and pick up a red can on a table. The can was modeled as a filled cylinder. The table was presented with two task space models which represent the space on the top board and between the top and second boards (Fig. 14), respectively; and all acceptable angles of the possible approach directions were determined as 45 [deg]. First, the stereo camera system equipped on the wrist of the manipulator recognized the position and orientation of the red can, and the system determined which task space model includes the can. By the geometric condition of the robot hand, the basic grasping configurations 1, 2, and 3 were selected. Then, the range of the angle relative to the center line of the cylinder, \u03c6, was restricted based on the approach direction depicted on the task space model of the table" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001114_0951192x.2010.528033-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001114_0951192x.2010.528033-Figure11-1.png", + "caption": "Figure 11. (a) JOB\u2019OT 12; (b) kinematic diagram of JOB\u2019OT 12.", + "texts": [ + "0000 0.6000 0.0000 0.4363 2.3562 3 R q3 0.0000 0.1300 1.5708 71.3090 1.0472 4 R q4 0.5500 0.0000 71.5708 73.3161 3.3161 5 R q5 70.0700 0.0000 1.5708 72.4435 2.4435 6 R q6 0.1000 0.0000 1.5708 73.1415 6.2831 Hitachi process robot (Figure 10) 5 0 \u2013 0.0000 0.0000 0.0000 0.0000 \u2013 \u2013 1 R q1 0.6500 0.0000 1.5708 71.0472 4.1888 2 R q2 0.0000 0.6000 0.0000 0.7854 2.4435 3 R q3 0.0000 0.8000 0.0000 72.3562 71.1345 4 R q4 0.0000 0.0000 1.5708 0.0000 3.1416 5 R q5 0.1000 0.0000 1.5708 71.6581 4.7997 JOB\u2019OT 12 (Figure 11) 6 0 \u2013 0.0000 0.0000 0.0000 71.5708 \u2013 \u2013 1 P 0.0000 q1 0.0000 1.5708 0.0000 4.1000 2 R q2 0.7000 0.0000 0.0000 72.8798 2.8798 3 P 1.5708 q3 0.0000 1.5708 0.0000 1.2000 4 P 1.5708 q4 0.0000 71.5708 1.2000 2.4000 5 R q5 0.0000 0.0000 1.5708 71.9199 1.9199 6 R q6 0.2300 0.0000 1.5708 71.9186 5.0615 Kuka IR 761/60.1 (Figure 12) 7 0 \u2013 0.0000 0.5000 0.0000 71.5708 \u2013 \u2013 1 P 0.0000 q1 0.0000 1.5708 0.8000 2.1300 2 R q2 0.9900 70.5000 71.5708 72.7925 2.7925 3 R q3 0.0000 1.4500 0.0000 72.8798 70.6109 4 R q4 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001974_09544097jrrt341-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001974_09544097jrrt341-Figure2-1.png", + "caption": "Fig. 2 ALE decomposition of the motion", + "texts": [ + " The velocity of the centre of the wheel is v0, and the wheel is rotating with a constant angular rolling velocity \u03c90 around a rigid axle of wheel T at point X 0. In the ALE method, x are Eulerian coordinates, and X are Lagrangian coordinates, and \u03c7 are ALE coordinates. The function x = \u03d5(X , t) maps the body from the initial configuration 0 to the current or spatial configuration , the function \u03c7 = \u03c7(X , t) maps the initial configuration 0 to the reference configuration domain \u0302, and x = \u03d5\u0302(\u03c7, t) maps the reference configuration domain \u0302 to the current or spatial configuration . The map of Lagrangian\u2013Eulerian and ALE domains is shown in Fig. 2. In the ALE method [13], the motion of a particle X at time t consists of a rigid rolling rotation described by \u03c7 = \u03c7(X , t) and a deform to point x described by x = \u03d5\u0302(\u03c7, t). \u03c7 is described as \u03c7 = RS (X \u2212 X 0) (1) where RS is the cornering rotation given by RS = exp(\u03c9t) and \u03c9 is the skew\u2013symmetric matrix associated with the rotation vector \u03c9 = \u03c9T . If the wheel is moved to the position y at the velocity v0 along the rail, so that y is described as y = x + v0t (2) The velocity of material particle of wheel in the relative kinematical description reads as v = y\u0307 = Dx Dt + v0 (3) where Dx Dt = \u2202\u03d5\u0302 \u2202t \u2223\u2223\u2223\u2223 \u03c7 + \u2202\u03d5\u0302 \u2202\u03c7 \u00b7 \u2202\u03c7 \u2202t (4) The circumferential direction S is defined by the following equation S = T \u00d7 (\u03c7 \u2212 X 0) R (5) whereT is the rigid axle of the wheel at X 0; R = |\u03c7 \u2212 X 0| is the radius of a point on the reference body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure8.13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure8.13-1.png", + "caption": "Fig. 8.13 The SCARA robot", + "texts": [ + " The computed torque control method allows online tuning of the position and velocity gains, as well as some changes in the controller\u2019s structure. For example, it is possible to turn on or off the compensation of spring torque and friction. In this way the student can see how important is the compensation of these two dynamic effects for good efficiency of the computed torque controller. After finishing the experiments, the students have to write a report with their results, where all methods are compared, evaluated and commented. The second experimental device is a two degrees of freedom robot of SCARA type (Fig. 8.13). The robot is without wrist, and so the motion is limited only to the X-Y plane. As actuators, two ESCAP 28D11 direct current motors mounted on the robot base are used. The advantage of this construction is that the moving masses, and therefore the inertias, are lower, so that higher accelerations can be achieved. For transfer of the motion to the robot joints, belt transmission combined with gear trains is applied. The gear ratio is the same for both joints, i.e., N = 105/28. Each of the motors is equipped with an incremental encoder for measuring the shaft velocity and the direction of rotation, as well as its position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000457_iros.2008.4651183-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000457_iros.2008.4651183-Figure1-1.png", + "caption": "Fig. 1. Robot\u2019s motion planning in wide open area", + "texts": [ + " Secondly, we address the travel cost in terms of collision likelihood along the path, which is not a concern in [13]. Our approach has similarity to recent work that incorporates exploration within SLAM framework, in particular, [3] and [14]. [14] also simulates SLAM for a given set of commands and chooses the one that maximizes information gain. However, the set of potential paths considered (the Voronoi diagram of the environment) is often small and the simulated SLAM is not integrated within an overall planning framework. For instance, when following a Voronoi diagram based path from A to B in Fig. 1 (shown as path1), the robot\u2019s localization uncertainty may accumulate quickly along the path (since the robot can not localize itself well due to absence of objects within the sensing range along the path), and the robot could get \u201clost\u201d and then remain lost as it reaches close to B and may collide with B. Our RRT-SLAM planner will return an alternative solution such as path2, which, although longer, is a better choice, since it allows the robot to localize itself when traveling along the path with its on-board range sensor, thereby reducing the chance of collision caused by the localization uncertainty when approaching the goal area near position B" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002245_1754337112442619-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002245_1754337112442619-Figure5-1.png", + "caption": "Figure 5. Constructed FE model of a grip.", + "texts": [ + "comDownloaded from A simulation model for a swing consists of an FE model of the grip corresponding to the lower lever of the robot and an FE model of the club. A swing for FEA was modelled by prescribing the positional coordinate data of three markers attached to the grip, which were obtained from the robot test, into the grip model. The FE model of the grip, which was constructed using four-node shell elements, consists of a hollow cylindrical body and three parallelograms, corresponding to the lower lever of the robot and markers attached to its surface, respectively, as shown in Figure 5. The positions of nodes at the centre of each parallelogram were prescribed such that they produced a swing comparable to that in the test. The grip was modelled as a simple shape with little regard for detailed construction of the geometry because the model requires it only to represent the swing profile rather than the geometry. The material models of the cylindrical body and parallelograms are linear elastic and rigid body, respectively. The material properties of the grip can be treated as a rigid body in view of the fact that the gripper of the robot has a significantly greater stiffness than the club shaft", + " The simulation results for the trajectories of the grip and the butt end of the shaft were obtained from the positional data of nodes corresponding to markers on the grip and that of the midpoint of nodes corresponding to markers on each side of the butt end, respectively. The simulation results were compared with the experimental data to confirm that the constructed grip model can accurately represent the swing motion. Figure 6 shows the time histories of the Euclidean distance between the grip nodes of the FEA and experiment from the beginning of the downswing to just before impact. Labels from A to C in the figure represent the nodes corresponding to the IDs of markers on the grip, as shown in Figure 5. The differences between the simulation and the experiment for all markers were accurate within 0.5mm. Figure 7 shows the time history of the Euclidean distance of the shaft butt between the FEA and the experiment from the beginning of the downswing to just before impact. The distance was accurate within 1mm, and the grip model provided the prescribed motion to the shaft. Meanwhile, the difference between the FEA and the experiment was larger than for the grip for the rigid body parts, and tended to have a random vibration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001764_s11036-009-0152-y-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001764_s11036-009-0152-y-Figure3-1.png", + "caption": "Fig. 3 (left) The robot\u2019s synchronization scheme. (center and right) A valid plan may still lead to an ICS during the next planning cycle", + "texts": [ + " While moving, the vehicles must avoid collisions both with obstacles and with other vehicles. Since the workspace is only partially observable, the vehicles must update their world model and state estimate given new sensory information. In parallel, they must compute in realtime trajectories towards their goals. To achieve this objective, a vehicle\u2019s function is broken down into a sequence of consecutive operational cycles. The various vehicle operations, shown in Fig. 2, are executed in a pipeline over these cycles (Fig. 3 left). For cycle (t : t + dt) vehicle Vi executes the following: 1. Up to time t: a localization and mapping routine updates the map Mi(t) and estimates future state xi(t + dt). 2. Given the map, a goal Gi(t) is computed for Vi. 3. Given Mi(t), Gi(t) a planner must compute a plan p(dt) before t + dt. 4. At time t + dt the plan p(dt) is executed at xi(t + dt). In this work we focus on step 3, i.e. on how to utilize communication so as to provide safety guarantees when multiple vehicles operate in close proximity", + " If two vehicles Vi, Vj at time t are not in collision with each other or with obstacles, then their corresponding states xi(t), x j(t) are compatible states: xi(t) x j(t). Two trajectories \u03c0i(xi(ti), pi(dti)) and \u03c0 j(x j(t j), pj(dt j)) are compatible trajectories (\u03c0i \u03c0 j) if the two trajectories do not cause collisions with workspace obstacles and: \u2200 t\u2032 \u2208[max{ti, t j} : min{ti + dti, t j + dt j}] : x\u03c0i(t\u2032) x\u03c0 j(t\u2032). Compatible trajectories between vehicles may still lead to an inevitable collision state from which a collision cannot be avoided in the future due to secondorder constraints [28]. A state xi(t) is an Inevitable Collision State (ICS) (Fig. 3 right) given the states {x1(t), . . . , xv(t)} if \u2200 \u03c0i(xi(t), pi(\u221e)): \u2203 (dt \u2227 j = i) so that \u2200 \u03c0 j(x j(t), pj(\u221e)) states x\u03c0i(dt) and x\u03c0 j(dt) are not compatible. 2.3 Problem definition Given the map Mi(t) and a state estimate xi(t + dt), the motion planning module of each vehicle Vi must compute before time (t + dt) a plan pi(dt\u2032) so that given the trajectories of all other vehicles \u03c0 j(x j(t + dt), pj(dt\u2032)) (\u2200 j = i): \u2022 \u03c0i(xi(t + dt), pi(dt\u2032)) \u03c0 j(x j(t + dt), pj(dt\u2032)) \u2022 State x\u03c0i i (dt\u2032) is not ICS", + " This is achieved by expanding a tree data structure (Tree) in the vehicle\u2019s state-time space using a sampling-based approach [18, 20, 21, 44]. From the expanded tree, a valid plan p(tn : tn+1) that results in the trajectory \u03c0(x(tn), p(tn : tn+1)) is selected. Algorithm 1 SAMPLING-BASED PLANNER Retain valid subset of from previous cycle while do Select a state on the existing Select valid plan given state Forward propagate trajectory if ( is not collision-free with obstacles) then Reject else Add to It is not sufficient for \u03c0(x(tn), p(tn : tn+1)) to be just collision-free, since it may lead to an ICS [28], as Fig. 3 (right) demonstrates. It is computationally intractable, however, to check if a state is truly ICS or not: all possible plans out of that state have to be examined to determine if there is an escape plan. It is sufficient, however, to take a conservative approach: if the vehicle can avoid collisions by executing a pre-specified \u201ccontingency\u201d plan \u03b3(\u00b7) out of a state x, then x is safe. In other words, state x is safe iff: \u2203 \u03b3(\u221e) s.t. \u03c0(x, \u03b3(\u221e)) is collision free. (2) In our experiments, the contingency plan we use for car-like vehicles is a breaking maneuver that brings the car to a complete stop as fast as possible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003168_cphc.201200772-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003168_cphc.201200772-Figure8-1.png", + "caption": "Figure 8. Setup for measuring the dynamics of wrinkling by light scattering. The sample is placed on a hot stage with transparent window, allowing the direct comparison of thermal and photoinduced wrinkling.", + "texts": [ + " All thermal depositions were performed with a Bestec customer-specified evaporation system. Silicon nitride as alternative cover layer was deposited on top of the azo layer by plasma-enhanced chemical vapor deposition.[12] The thickness of all layers was confirmed by spectroscopic ellipsometry (Woollam VASE). Typical thicknesses are ~120\u2013150 nm of AZOPD, and 30\u2013 50 nm for the capping layer. The kinetic evolution of the wrinkles was measured by time-resolved light scattering with the setup given in Figure 8. The samples were placed on a programmable heating stage (Linkam THMS 600) with a hole in the silver support to allow thermally induced wrinkling as well as optically induced wrinkling by irradiation with a laser beam (Elforlight B4, 473 nm, expanded to a diameter of 2 mm). The intensity of irradiation was controlled by a neutral density filter set and measured with a calibrated photodiode (OphirNova II). Diffraction patterns were recorded in backscattering geometry by diffraction of a red helium neon laser beam (Uniphase 1124 m, 632" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000465_iet-smt:20060018-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000465_iet-smt:20060018-Figure3-1.png", + "caption": "Fig. 3 Eigenmode 4", + "texts": [ + " Inspection of the results for mode 2 in Table 1 leads to the conclusion that, for this mode, the contributions of the magnetic forces and the magnetostriction to the vibrations are in phase and thus add up. Fig. 1 Magnetostriction measurements IET Sci. Meas. Technol., Vol. 1, No. 1, January 200722 The four-pole IM has the same characteristics as the two-pole IM. Table 2 shows the results of the eigenmode analysis for the most important eigenmodes of the four-pole machine for a frequency of 100 Hz. In the case of the four-pole machine, it is clear that the fourth-order modes, that is modes 4 and 40, are the most important modes. In Fig. 3, vibration mode 4 is shown. The contour of the undeformed machine hull is represented by the black line. Mode 40 has the same shape as mode 4, but is shifted over 22.58. The generalised vibrations of modes 4 and 40 at 100 Hz have a phase shift of 908. As explained in the preceding paragraph, the combination of these modes produces rotating vibration waves. Inspection of the results for mode 4 in Table 2 leads to the conclusion that, for this mode, the contributions of the magnetic forces and the magnetostriction to the vibrations are in counter phase and thus are subtracted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003028_bf02326512-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003028_bf02326512-Figure1-1.png", + "caption": "Fig. 1--Segment of a curved bar", + "texts": [ + " Application of the Method of Internal Constraints to the Study of the Vibration of a Toroid The procedure used in the application of the method of internal constraints for deriving the equations of mot ion for a one-dimensional elongated solid with a curvilinear baricentric axis is now briefly described. The elongated solid is referred to a curvilinear left-handed coordinate system ( O X Y Z ) in which the origin 0 coincides with a generic point of the central axis of the bar, X coincides with the central axis (and has as positive direction the positive direction of the central axis), Y coincides with the principal normal and is directed positively toward the center of curvature, and Z coincides with the binormal at 0 (see Fig. 1). The expression of the square of the line element ds2 in the curvilinear system O X Y Z is ~I: ds ~ = dx2[(1 -- yc)~ + (yr) ~ + (zr) 2] + dy ~ + dz ~ + 2 y r d x d z - 2zTdxdy (1) where c and r represent, respectively, the curvature and torsion of the central axis. In the particular case of t he to ro id , T = 0 and c = Co = constant and eq(1) becomes: ds ~ = dx~[1 - YCo] ~ + dY ~ \"4- dz ~ In this case, the fundamental tensor is expressed by the metric: * an = (1 - y c o ) ~ a12 = 0 a13 = 0 a21 = 0 a22 = 1 a13 = 0 (2) a3~ = 0 a~2 = 0 a33 = 1 a = (1 -- yco) 2 ~r = (1 --yc0) Expressing the components of the elastic displacement u, v, w in a Taylor ' s series expansion in terms of the y and z coordinates respectively in the x, y, and z directions, the following expressions will be used: " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001565_iemdc.2009.5075386-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001565_iemdc.2009.5075386-Figure3-1.png", + "caption": "Figure 3: Stator winding with induced fault", + "texts": [ + " The supply to the motor is generated from a 250kW synchronous generator driven by a 250kW DC motor, whereby voltage and frequency may be varied. Since the induction motor does not allow direct access to the stator windings, an alternative method of inter-turn fault implementation is used. It is noted that an inter-turn fault causes large circulating currents (If) to flow within the shorted coil, which increases with increasing fault severity. Thus, in order to reproduce inter-turn fault conditions, a variable resistor (Rf) was place in parallel with a stator coil as shown in Fig. 3. By varying the resistance, the severity of the inter-turn fault could be varied. Four conditions were examined, ie. the machine in it\u2019s original state and the machine with 0.5%, 1% and 1.5% shorted turns. The first set of results highlights the key characteristics of the EPVA technique when subject to changes in load, speed (Voltage / Frequency control) and fault severity. The next section proposes an improvement to the EPVA technique by incorporating wavelets, whereby key findings will be discussed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003583_iros.2011.6094491-FigureI-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003583_iros.2011.6094491-FigureI-1.png", + "caption": "Fig. I. In momentum control, when the ground reaction force and the center of pressure computed from desired momentum rate change are not simul taneously achievable, two distinct scenarios can arise. (a) Fully respecting the desired linear momentum while compromising angular momentum (if needed) results in a postural balance control without stepping. (b) Contrarily, fully respecting the desired angular momentum while compromising linear momentum (if needed) results in a stepping motion.", + "texts": [ + " For traditional balance controllers([2], [3]), the focus has been primarily on linear momentum which is a scaled version of the CoM velocity. As attention has been given to angular momentum and the important role it plays in balance ([4], [5], [6]), the presented controller enables us to control both the angular momentum and linear momentum of a humanoid. Our balance controller in [I] attempted to attain the desired linear momentum at the cost of sacrificing precise control of the angular momentum. This encourages the robot to maintain balance by making postural adjustments without stepping (See Fig. I(a)). In this paper we show the opposite approach (See Fig. I(b)) in which the controller prioritizes angular momentum over linear momentum. Combined with a stepping controller for the swing leg, this leads to a natural looking reactive stepping such that a humanoid can recover from a strong push by taking a step. Seung-kook Yun and Ambarish Goswami are with Honda Research Institute US., 425 National Ave, Suite 100, Mountain View, CA, USA syun@honda-ri.com, agoswami@honda-ri.com When a robot is under external disturbance, finding an ap propriate stepping location for balance recovery is important to minimize control effort", + " When the same push is applied to the robot with the postural balance controller, the robot fails to take a step because it sacrifices angular momentum too much which must accompany large change of the upper body orientation. The controller tries to keep the CoM close to the reference trajectory and causes the bending motion of the upper body, which results in joint limit violation of the hip. Note that we use the exactly same trajectories for the postural balance controller (linear momentum respecting) and the reactive stepping controller (angular momentum respecting). Fig. II shows the rate changes of linear and angular mo mentum (control inputs) from the simulations by the postural balance controller and the reactive stepping controller. In the lower left of Fig. 11, the large gap between the desired rate change of angular momentum and the admissible rate change shows that the postural balance controller tries to track the linear momentum by sacrificing precise control of the angular momentum. On the other hand, in the right of Fig. 11, the reactive stepping controller generates the admissible rate change of angular momentum which is closer to the desired rate, by allowing the tracking error of the linear momentum" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000180_cca.2008.4629582-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000180_cca.2008.4629582-Figure1-1.png", + "caption": "Fig. 1. Model of quadrotor", + "texts": [ + " The procedure might even be called \u2018forward stepping\u2019, in contrast with the well-known backstepping method. Section 2 of this paper discusses the basic quadrotor dynamics which is used for control law formulation. Section 3 shows dynamic inversion of a nonlinear state-space model of a quadrotor. Sections 4 discuss the robust control method to stabilize the internal dynamics. In the final section simulation results are shown to validate the control law discussed in this paper. II. QUADROTOR DYNAMICS Fig. 1 shows a basic model of a unmanned quadrotor. The quadrotor has some basic advantage over the conventional helicopter. Given that the front and the rear motors rotate counter-clockwise while the other two rotate clockwise, 978-1-4244-2223-4/08/$25.00 \u00a92008 IEEE. 1189 gyroscopic effects and aerodynamic torques tend to cancel in trimmed flight. In this subsection we will describe the basic state-space model of the quadrotor. The dynamics of the four rotors are relatively much faster than the main system and thus neglected in our case" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002914_1350650113511960-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002914_1350650113511960-Figure3-1.png", + "caption": "Figure 3. Textured surface with slip and no-slip region.", + "texts": [ + " d xd z no-slipzone 8>>>>< >>>>: \u00f023\u00de Present model is based on thin film lubrication, which predicts 2D pressure field p(x, z) between the mating surfaces and does not facilitate to impose slip condition on the vertical surface of the textures. Therefore, a more complex 3D hydrodynamic model may be used to predict the effect of vertical surface of the textures. Present work considers complex slip surface pattern, where no-slip condition is ensured in one region of specific geometry (top of the textures) with the use of high critical threshold shear stress; whereas slip condition is imposed on the other region (around the textures) with a lower value of critical threshold shear stress, as shown in Figure 3. To validate the numerical solution, the present model is reduced to a model described by Salant and Fortier11 and the comparison of result is shown in Figure 4, where good agreement is achieved. The numerical results without slip for textured parallel sliding contacts with different shape of textures are also validated with Siripuram and Stephens.21 A good correlation in the comparison of film thickness (above protrusion) is obtained as shown in Figure 5. Slight variation observed, may be due to difference in mesh size" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000031_6.2008-7413-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000031_6.2008-7413-Figure2-1.png", + "caption": "Figure 2. Central ring (left), vertical connecting pipes (middle), and complete body frame (right).", + "texts": [ + " More precisely, the frame is a light-weight construction consisting of a symmetrical (both horizontally and vertically) top and bottom frame. These two partial frames, as well as the eight vertical connecting pipes, are built entirely from CF and AL. The rotor axles, which connect the top and bottom frame at their far ends, are manufactured from TI with the aim to ensure high cohesiveness at reduced weight. The frame design of the JAviator not only incorporates light-weight materials, it also provides high mechanical integrity. Figure 2 (left) shows the bottom center ring with four AL flanges installed that connect to the side-arm pipes. These flanges are designed to also hold eight vertical CF pipes, as shown in Figure 2 (middle). The top ring, which is identical to the bottom ring and similarly equipped with four AL flanges, completes the JAviator\u2019s body frame, depicted in Figure 2 (right). In this illustration, the bottom ring is already closed with a thin CF plate to carry the onboard battery. An equally shaped top plate, not assembled in the present figure, is used to mount the inertial measurement unit (IMU). The body of the JAviator is formed like a cage, which is simple in its design, features fast and easy (re)assembly, and also ensures high integrity at relatively low weight. Compared to most other quadrotor bodies, this cage design acts like a protection frame for the onboard electronics and has proven in crashes to withstand collisions without serious damage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000366_elan.200804278-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000366_elan.200804278-Figure3-1.png", + "caption": "Fig. 3. Cyclic voltammograms of a) CPE, b) as (a) in presence 0.1 mM nitrite, c) CPE/CuHCF in absence of nitrite, d) CPE/ CuHCF in solution as (b), v\u00bc 10 mV s 1.", + "texts": [ + " This study showed that the height of the cyclic voltammograms decreases in basic medium (pH> 7), indicating a gradual dissolution of the CuHCF film according to a hydrolysis reaction as follows: Fe(CN)6 4 \u00fe 2OH !Fe(OH)2\u00fe 6CN 2Fe(OH)2\u00feH2O\u00fe 1/2O2! 2Fe(OH)3 Therefore, the utility of the modified electrode is limited to neutral and acidic solutions. Due to their high electrochemical reversibility, MHCFs are widely used in electrochemistry as electron mediators to shuttle electrons between analytes and substrate electrodes [29, 38 \u2013 41]. In this work, nitrite is chosen as a model to elucidate the electrocatalytic ability of CPE/CuHCF. Shown in Figure 3 are the cyclic voltammetric responses of CPE in the absence (curve a) and in the presence of 0.1 mM nitrite (curve b) in a 0.1 M KNO3\u00fe 1 M HNO3 as supporting electrolyte. From the curve b, we can find no obvious electroreduction of nitrite on a bare electrode between 0.0 and 1.0 V. On the other hand, for the CPE/ CuHCF (compare curves c and d), cathodic current was found to increase at 0.55 V in the presence of nitrite, indicating that CuHCF films can electrocatalyze the reduction of nitrite (ECJ mechanism; Scheme 1) The reduction of nitrite in the presence of CuHCF film at the surface of CPE occurs at a potential about 800 mV more positive than that the absence of film" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002027_j.procs.2010.04.035-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002027_j.procs.2010.04.035-Figure4-1.png", + "caption": "Figure 4: SeT laboratory electrical vehicle (left), vehicle architecture (right)", + "texts": [ + " This vector is then applied as a command law to the vehicle. 322 F. Gechter et al. / Procedia Computer Science 1 (2010) 317\u2013325 / Procedia Computer Science 00 (2010) 1\u20139 7 The model described in the previous section has been implemented thanks to the MadKit multiagent platform2. Experimentation are used to validate some model characteristics. The validation of the proposed reactive multi-agent system has been realized on real vehicle. The System and Transportation Laboratory has two electric vehicles (cf. figure 4 on the left). These vehicles have been automated and can be controlled by an onboard system. Figure 4 shows the vehicle system onboard. An onboard computer receives direction and speed instructions from the driver (numerical steering wheel) and sends direction and speed intructions to the vehicle control. The reactive multi-agent system has been implemented on this onboard computer. The next subsection proposed to validate the Car-driving assistance in two cases: simple or several obstacles on the road. The first experiment is to evaluate the path taken by the vehicle, in the case of the appearance of a single obstacle (figure 5 on the left)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000152_0022-4898(81)90016-1-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000152_0022-4898(81)90016-1-Figure1-1.png", + "caption": "FIG. 1. (a) Wheel with slip angle at zero camber angle (plan view of wheel). (b) Wheel", + "texts": [ + " What is required is a model o f the tyre-soil interaction which will allow the forces on a steered wheel to be readily predicted f rom a knowledge o f the wheel parameters and soil condition. This paper describes such a model for an undriven wheel. DIMENSIONAL ANALYSIS The performance o f bo th driven and undriven wheels at zero slip angle has been modelled with some success using dimensional analysis [6--9]. It was therefore decided to perform a similar analysis for undriven steered wheels. The parameters required to describe the system are listed in Table 1 and the forces acting on the wheel shown in Fig. 1. 28 D. GEE-CLOUGH and M. S. SOMMER For typical agricultural operations, forward speed V will be low and soil strain-rate effects would also be expected to be low. The groups z Vd/W and V2/gd were therefore ignored. The ratio hid has been shown to have little effect on tyre performance at zero slip angle and this group was also ignored as was the effect of wheel skid. The experiments were carried out with camber angle equal to zero. In the soil conditions used in the experiments it was found that soil cone index value varied by about 6 : 1, cohesion by 4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001957_j.protcy.2012.03.012-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001957_j.protcy.2012.03.012-Figure1-1.png", + "caption": "Fig. 1. Wiimote\u2019s degrees of freedom", + "texts": [ + " Similar projects like the one proposed herein are WiiGee [13], GesApp [14], part of the Google Summer of Code 2008, and [15] have been developed to evaluate and extract features of trajectories and movements taken from accelerometers. A gesture is known as the data obtained from a movement made with a HID. The gesture taken from a Wii remote is represented by characteristic patterns of incoming signal data, i.e. vectors representing the current acceleration of the controller in all three dimensions [13]. A Wii remote (a.k.a.Wiimote) has six degrees of freedom (as shown in Fig. 1); this means it can move along the X, Y and Z axes, as well as it can rotate on them (pitch, roll and yaw). To measure the acceleration in the Wiimote a ADXL330 3-axis accelerometer is used. It provides a +/ \u2212 3g 10% resolution three axes linear accelerometer produced by Analog Devices c\u00a9. The Wiimote has a high sampling frequency; it can deliver up to 100 values per second for each of the accelerations in X, Y and Z [16]. If the Wiimote were in free fall with the Z axis perpendicular to the floor, it would return values very close to 0 for the acceleration in Z" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002041_ajpa.21537-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002041_ajpa.21537-Figure1-1.png", + "caption": "Fig. 1. (A) Two-dimensional musculoskeletal model of the Japanese macaque consisting of nine rigid links and eight principal muscles. See Table 2 for muscle numbers. (B) Model of contact between the floor and foot.", + "texts": [ + " As a first step, therefore, we constructed a 2D musculoskeletal model of the bipedal Japanese macaque consisting of nine links representing the HAT (head, arms, and trunk), thighs, shanks, and feet that are represented by two parts: a tarsometatarsal part modeled by a rigid body; and a mass-less phalangeal part (hence the dynamics were not considered), based on our recently constructed anatomically based whole-body musculoskeletal model (Ogihara et al., 2009). Here, we considered eight principal muscle groups classified according to muscle disposition (Fig. 1A). Dimensions and inertial parameters of the limb segments and muscle parameters were determined based on this 3D model (Tables 1 and 2). Contact between the floor and foot was modeled by five vertical viscoelastic elements and five horizontal viscous elements attached to the sole and phalangeal part (Fig. 1B). The metatarso- Note: Center of mass (COM) is represented as a fraction of segment length from the proximal end. MofI 5 moment of inertia about the COM. Parameters were determined based on the anatomically based whole-body musculoskeletal model of the Japanese macaque (Ogihara et al., 2009). American Journal of Physical Anthropology phalangeal joint was assumed to be kinematically prescribed (predetermined based on measured kinematic data), as the mass and inertia of the phalangeal part are very small, making the simulation calculation particularly difficult if we consider the dynamics of this segment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002483_00207721.2012.669869-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002483_00207721.2012.669869-Figure1-1.png", + "caption": "Figure 1. Group of two-link robots. (a) Group topology; (b) robot model (Chung 2007).", + "texts": [ + " Taking the derivative of Va in (78) with respect to the trajectories of system (77), and choosing ui\u00bc u\u0302i\u00fe ~ui, u\u0302 as the nominal control law (14) or (15), and ~ui as the redesign (82), we have _Va k3k \u00f0L Im\u00deqk 2 2 k4z Tz k Xn i\u00bc1 ikzik 2k ik 2\u00fe k Xn i\u00bc1 zTi i ~ i\u00fe k Xn i\u00bc1 ~ Ti 1i _~ i \u00f083\u00de for some k3, k4> 0. Still choosing _~ i \u00bc i T i zi, we have _Va k3k \u00f0L Im\u00deqk 2 2 k4z Tz k Xn i\u00bc1 ikzik 2k ik 2 0: \u00f084\u00de Since vd and k ik are bounded, we apply the invariance-like theorem (Theorem 8.4 in Khalil (2002)) and conclude that k \u00f0L Im\u00deqk 2 2! 0 and zT z! 0 as t!1, which is equivalent to qi ! qj, _qi ! vd as t!1, 8i, j2I . The damping term izik ik 2 2 actually fortify the adaptive process and smoothen the transient trajectories. We will show this point in the simulation. As shown in Figure 1, assume the group has four twolink manipulators and the communication topology is a digraph. For convenience, we set the link weights aij\u00bc 1, j2Ni, i, j2 {1, 2, 3, 4}. As shown in Figure 1(b), the two-link manipulator model has two degrees-of-freedom, the generalised coordinates q\u00bc [ 1, 2] T. The moment of inertia about the centre of mass of each manipulator is given as I1 and I2. The lengths of each manipulator are l1 and l2, while lc1 and lc2 denote the distance from the previous D ow nl oa de d by [ N or th C ar ol in a St at e U ni ve rs ity ] at 0 6: 49 1 5 D ec em be r 20 12 joint to the centre of mass of the next link. Also, m1 and m2 denote the mass of each manipulator link, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001956_iros.2010.5651948-Figure10-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001956_iros.2010.5651948-Figure10-1.png", + "caption": "Fig. 10. 3D representation of the optimal points visit order solution", + "texts": [ + " In fact, in the other generations the convergence rate is less than 15%. In fact the mutation rate of 0.1 give the best average of 11.78 sec, while the mutation rate of 0.4 gives 13.51sec of the task time. The optimization of our GAs algorithm (with generation number of 500000) leads to find an optimal time of 6.12 sec. The CPU time to find the optimal solution was 13min and 5 sec. This solution was found in the iteration number of 222857 with the following chromosome: 12, 11, 8, 9, 10, 7, 6, 5,4,2,1,3. Fig. 10 represents the final solution found by the proposed algorithm. In this figure only the visited points showed. Is we cannot predict the robot position and or the configurations of IKM of the robot. This visit order is corresponding to the configurations of 3,3,3,3,3,4,4,3,3,3,3,3 corresponding to the placement of 1,7,5 on x, y and z respectively. In this paper we have proposed and tested a novel optimization approach for task execution time of industrial robots based on Genetic Algorithms. Our approach takes into consideration the important factors which can effects the task\u2019s time which are multiple configurations of the IKM of the robot and the relative placement between the robot and the task points as well as the order of visited points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001190_icelmach.2010.5608143-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001190_icelmach.2010.5608143-Figure9-1.png", + "caption": "Fig. 9 Overview of the prototype magnetic gear.", + "texts": [ + "029 Nm, respectively, which are very small compared with its maximum transmission torque. This implies the flux path helps to reduce the cogging torque on the high-speed rotor. The orders shown in Table I are found. The orders not shown there are the harmonic components (H.C.) by its pole pairs and stationary pole pieces. These orders appear because of the error of the mesh in the analysis model. VI. VERIFICATION BY A PROTOTYPE A prototype based on the specifications mentioned above was manufactured, and is shown in Fig. 9. This prototype is set to the torque measuring system shown in Fig. 10, and the constant rotation speed was given to both rotors in accordance with the gear ratio by the AC servo motors. A rotation speed of 0.5 rpm was given to the high-speed rotor to ignore the effect of the eddy current. The synchronous transmission torque shown in Fig. 11 was measured. The orders contained in the cogging torque on both rotors are shown in Figs. 12 and 13. The cogging torque on the high-speed rotor and lowspeed rotor is 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002343_s10846-011-9583-3-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002343_s10846-011-9583-3-Figure2-1.png", + "caption": "Fig. 2 Schema of the four-rotor rotorcraft", + "texts": [ + " The relation between the frame {RB} and the frame {RE} can be obtained by the position vector \u03be = (X, Y, Z )T \u2208 3, which describes the displacement of the center of mass G in the inertial frame and the orientation vector \u03b7 = (\u03c8, \u03b8, \u03c6)T \u2208 3 where (\u03c8, \u03b8, \u03c6) are the three Euler angles (yaw, pitch, and roll) representing the orientation (Fig. 1). The generalized coordinates for the rotorcraft are q = (X, Y, Z , \u03c8, \u03b8, \u03c6)T \u2208 6 (1) The dynamic model of the rotorcraft can be developed using the Lagrangian form of the dynamics [2, 17, 18]. A simplified model can be obtained by representing the quad-rotor rotorcraft as a solid body evolving in 3D space and subject to one force and three moments (Fig. 2). The translational kinetic energy of the rotorcraft can be described as Ttrans = 1 2 m\u03be\u0307T \u03be\u0307 (2) where m is the mass of rotorcraft. The rotational kinetic energy is given by Trot = 1 2 \u03b7\u0307T J\u03b7\u0307 (3) where J = diag ( Ixx, Iyy, Izz ) is the inertia matrix expressed in terms of the coordinates \u03b7. The only potential energy that needs to be considered is caused by the gravitational force U = mgZ (4) In such a case the Lagrangian is given by L (q, q\u0307) = Ttrans + Trot \u2212 U (5) L (q, q\u0307) = 1 2 m\u03be\u0307T \u03be\u0307 + 1 2 \u03b7\u0307T J\u03b7\u0307 \u2212 mgZ (6) Then the model for the quad-rotor helicopter dynamics can be deduced by using the LagrangeEuler equations with external generalized force d dt \u2202L \u2202q\u0307 \u2212 \u2202L \u2202q = F (7) where F = ( F\u03be , \u03c4 ) " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002455_lpor.201300078-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002455_lpor.201300078-Figure4-1.png", + "caption": "Figure 4 (a) The optical layout for the characterization of the optical power imposed by the cylindrical surface. (b) A picture of the optical bench, showing the impinging beam, the reflected and the 1st diffracted order.", + "texts": [ + " It should be noted that this is rigorously true regardless of the grooves\u2019 orientation over the surface, thanks to the perfect Euclidean geometry of cylindrical surfaces. The astigmatism of the system can be exploited either to compensate for aberrations, or in connection with other cylindrical surfaces with optical power in the orthogonal direction to give a spherical shape to the wave front. The experiments were performed by sticking a SCBI stretchable grating onto the surface of a cylinder of radius R = 164 mm machined from an aluminum piece with no optical finishing of the surfaces. Figure 4a shows a sketch of the optical layout, with the grooves of the grating in the vertical direction, parallel to the rotation axis of the cylindrical surface. The plane wave front of a collimated laser beam impinges onto the surface of the grating, from which it is diffracted. In Fig. 4b a picture of the impinging, reflected and diffracted beams is shown, where the beam paths are evidenced by light scattered from vapors of liquid nitrogen. Due to the cylindrical shape of the surface, the system is endowed with a pure astigmatic optical power in the tangential direction, the curvature being vanishing in the other, so that the sagittal focal length is infinite (see Fig. 4a). The 0th and 1st diffraction orders have been collected with a CCD camera along the corresponding propagation directions, and the intensity distributions recorded at several distances from the grating. The sizes of the diffracted spots have been measured and the divergence provided a measure of the tangential focal length of the system. These results are compared to that obtained from basic optics: with R the cylinder radius, \u03b1 and \u03b2 the incidence and diffraction angles, the tangential focal length is given by [31, 43]: f = R cos2 \u03b2 cos \u03b1 + cos \u03b2 (2) Note that, due to the finite thickness of the gratings, the deformation determines a slight change in the grooves\u2019 spacing, from 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000996_j.jsv.2009.11.014-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000996_j.jsv.2009.11.014-Figure1-1.png", + "caption": "Fig. 1. Geometry of the beam.", + "texts": [ + " It is assumed that the locus of the cross-sectional centroid of the beam is a continuum curve l in space. The tangential, normal and bi-normal unit vectors of the curve are t, n and b, respectively. The Frenet\u2013Serret formulae, for a smooth curve, are [1]: t0 \u00bc k1n; n0 \u00bc k1t\u00fek2b; b0 \u00bc k2n; (1) where superscript prime denotes the derivative with respect to s. s, k1 and k2 are arc coordinate, curvature and torsion of the curve, respectively. We introduce x and Z directions in coincidence with the principal axes through the centroid O1, as shown in Fig. 1. The angle between the x axis and normal n is represented by y, which is generally a function of s. If the unit vectors of O1x and O1Z are represented by ix and iZ, then ix \u00bc n cosy\u00feb siny; iZ \u00bc n siny\u00feb cosy: (2) From Eqs. (1) the following expressions are obtained: t0 \u00bc kZix kxiZ; ix 0 \u00bc kZt\u00feksiZ; iZ 0 \u00bc kxt ksix; (3) where kx \u00bc k1 siny; kZ \u00bc k1 cosy; ks \u00bc k2\u00fey0: (4) Based on the assumption that the cross sections of the beam do not deform in its own plane, but is free to warp out of the plane, the dynamic displacement of the beam consisting of stretching, bending and torsion is expressed by u\u00bcWt\u00feUix\u00feViZ; (5) in which W \u00bc us\u00f0s; t\u00de\u00feZjx\u00f0s; t\u00de xjZ\u00f0s; t\u00de\u00fea\u00f0s; t\u00dew\u00f0x;Z\u00de; U \u00bc ux\u00f0s; t\u00de Zjs\u00f0s; t\u00de; V \u00bc uZ\u00f0s; t\u00de\u00fexjs\u00f0s; t\u00de; (6) where three displacement components of the cross section in the s, x and Z directions are fully represented by six rigid body modes, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003079_msf.723.332-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003079_msf.723.332-Figure4-1.png", + "caption": "Fig. 4 Trochoidal tool path.", + "texts": [ + " The experimental setup is shown in Fig.2, the machine tool used is YH-VMC850L four axis machining center, and the cutter used is carbide cutting tools with four evenly spaced teeth, 8mm in diameter. The force measure device adopted is Kistler 9255B dynamometer. The workpiece material used here is Nickel-based superalloy used in aero engine. In the machining process, a slot with 0.5mm in depth, 20mm in width and 30mm in length, as shown in Fig.3. The trochoidal tool path and its parameters are shown in Fig.4. As for contrast, a normal slot machining strategy is also presented, the cutter first create a 8mm width slot with slot milling and the material in both sides are then removed with straight tool path. The cutting parameters are listed below: Slot milling: Spindle speed: 2388rpm; feed rate: 100mm/min Trochoidal milling: Spindle speed: 2388rpm; feed rate: 1570mm/min In order to get the same machining time, the feed rate in trochoidal milling is set to be much larger than slot milling. The machining experiments are executed with two separate new cutters, and the same cool condition and coolant is used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000466_iros.2008.4650802-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000466_iros.2008.4650802-Figure14-1.png", + "caption": "Fig. 14. Production process of half hitch.", + "texts": [ + " The description of the intersections for the overhand knot production process can be represented by the following: El \u2212 Er (Fig. 12(d)) El \u2212 C\u0302\u2212 1 \u2212 C\u0302\u2212 1 \u2212 Er (Fig. 13(a)) El \u2212 C\u0302\u2212 1 \u2212 C\u2212 2 \u2212 C+ 3 \u2212 C\u0302\u2212 1 \u2212 C\u2212 2 \u2212 C+ 3 \u2212 Er (Fig. 13(b)) El \u2212 C\u0302\u2212 1 \u2212 C\u2212 2 \u2212 C\u2212 3 \u2212 C\u0302\u2212 1 \u2212 C\u2212 2 \u2212 C\u2212 3 \u2212 Er (Fig. 13(c)) The eight knot and the stevedore\u2019s knot can be produced based on skill synthesis in the same way. In this section, we consider the knotting process of a knot generated by one rope and one object. As an example, we analyze a \u201chalf hitch\u201d. Half hitch (Fig. 14) The half hitch is one of the knots that make a connection between a rope and an object. Although it is an easy task to make this knot, the strength of the knot is very low. However, the strength can be increased by combining the half hitch with other knots. Production process of half hitch Here, we omit the intersection description of the half hitch. The left end and the right end of the rope are represented by l1 and l2, and the left end and the right end of the object are represented by r1 and r2. The description of intersections on the rope and the object is performed in the order of initial location. First, the intersection C+ 1 is created by rope permutation (Fig. 14(b), (c)). Second, the rope is wrapped around the stick by rope moving to produce the intersection C+ 2 (Fig. 14(d)). Next, the intersection C\u0302\u2212 1 is made by loop production (Fig. 14(e)). Finally, the half hitch is finished by performing rope permutation twice and rope pulling once (Fig. 14(f), (g)). A knot generated by two ropes can be considered in the same way; however, it is omitted in this paper. The experimental system consists of a high-speed multifingered hand, high-speed tactile sensors, and a high-speed visual sensor. The hand has three fingers and two wrist joints. The joints of the hand can be closed at a speed of 180 deg./0.1 s. The tactile sensor measures the center position of a twodimensionally distributed load, and the total load is measured within 1 ms. This sensor is used for grasp force control during rope permutation [1]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003472_j.cirpj.2011.01.010-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003472_j.cirpj.2011.01.010-Figure7-1.png", + "caption": "Fig. 7. Scheme of the simulation model.", + "texts": [ + " The preferred regime for powertrain bearings is the hydrodynamic lubrication regime because it generates the least friction and wear on the sliding partners. The aim of surface structuring is twofold: Reduce the friction coefficient along the three regimes. Shift the Stribeck curve toward the left in order to overcome the boundary and mixed lubrication regimes at lower operating velocities. The ability of surface microstructures to increase hydrodynamic pressure and the contribution of their geometric parameters to reach this goal were studied using the Computational Fluid Dynamics (CFD) package COMSOL Multiphysics. The Multiphysics model (Fig. 7) is based on the incompressible Navier\u2013Stokes equation (1). A lubrication gap initially set at s = 20 mm exists between the static, microstructured base body and smooth counter body. The relative speed of the parallel surfaces was kept at v = 1 m/s. The intermediary substance is oil with a viscosity h of 0.79 Pa s, a density r of 887.68 kg/m3 and a temperature of 293 K. r\u00f0vr\u00dev \u00bc r\u00bd pI \u00fe h\u00f0rv \u00fe \u00f0rv\u00deT\u00de \u00fe f (1) where v is the fluid velocity, p is the pressure, I the identity matrix, (5v)T the transpose of (5v) and f the body forces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure3-1.png", + "caption": "Fig. 3. Functional\u2013geometrical scheme of the gear with \u201ctranslating wheel \u2013 eccentric sector\u201d: a. central reference position, b. current position.", + "texts": [ + " 2, carries out a constant transmission ratio. For such gearbox, an increase in the transmission ratio by 30% from the middle position towards the ends of the nut can be satisfactory, and this can be realized with the solutions shown in Fig. 1b,c. In this way, the mechanical steering gearbox can create the effect/feeling of an assisted steering system. In accordance with the intended application, the geometry of such a gear will be correlated with the construction of a steering box of type \u201cscrew \u2013 nut rack \u2013 sector\u201d (Fig. 3a). In this application, the rotation of the steering wheel is transmitted to the screw 1, which produces the right/left movement of the geared nut 1\u2032 in contact with the segment (sector) 2. In the current position (Fig. 3b) there have been marked the following geometrical & kinematical parameters: a \u00bc O1O2 \u2013 fictive center distance (of angle \u03b1=\u2220AO1C=\u2220EO2C), \u03b1w=\u2220AO1y=\u2220EO2y \u2013 current pressure angle, \u03b10=\u2220A0O1y=\u2220E0O2Or \u2013 reference pressure angle (samewith that in the central position), r2r \u00bc OrCr \u2013 centroid radius of the geared segment, \u03b3=\u2220(O1O2,O1 0O2 0) \u2013 angle between the centers lines. It is found that the kinematical\u2013geometrical parameters depend on the current position of the wheel, meaning the displacement S1of the rack wheel and the rotation \u03c62 of geared segment. In these terms (Fig. 3b, corresponding to Fig. 1b), a \u00bc O1O2 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi S1 \u00fe e sin\u03c62\u00f0 \u00de2 \u00fe a0 \u00fe e cos\u03c62 2q ; a0 \u00bc a0\u2212e \u00bc m0 2 z1 \u00fe z2\u00f0 \u00de\u2212e; \u00f03\u00de \u03b1 \u00bc arccos rb1 \u00fe rb2 a ;\u03b1w \u00bc \u03b1 \u00fe \u03b3; rb1;2 \u00bc m0 2 z1;2cos\u03b10;\u03b10 \u00bc 200; \u00f04\u00de \u03b3 \u00bc arctg S1 \u00fe esin\u03c62 a0 \u00fe ecos\u03c62 ; \u00f05\u00de r2r \u00bc OrCr \u00bc OrO 0 1\u2212O0 1Cr \u00bc a0\u2212 S1tg\u03c8\u00fe rb1\u2212 S1 cos \u03c8 1 sin \u03c8 ;\u03c8 \u00bc 90\u2218\u2212\u03b1w: \u00f06\u00de From a kinematic point of view, teeth contact on the line of action means rolling without sliding on the base circles, the line AE rolling of the circle of radius rb1, and rolling up on the circle of radius rb2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000560_00207170801930217-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000560_00207170801930217-Figure1-1.png", + "caption": "Figure 1. Underactuated rigid spacecraft.", + "texts": [ + " The globally asymptotic stability of the closed-loop systems of both axisymmetric spacecraft and asymmetric spacecraft are also analysed in this section by using LaSalle\u2019s theorem (Khalil 1996). To demonstrate the efficacy of the proposed control algorithms, the simulation results are presented in Section 4. The concluding remarks are given in Section 5. 2. Problem statement Introduce two coordinate frames: the inertial coordinate frame denoted by frame Oxyz, the body fixed coordinate frame denoted by frame Sb1b2b3 . Consider the system describing the rotation of the spacecraft around its centroid as shown in Figure 1, where S is the mass centre of the spacecraft; ~v is the fixed unit vector in the frame Oxyz; ~b1, ~b2, ~b3 are the principal axes of the spacecraft with no thrusters on the third axis ~b3, i.e., the third axis is underactuated. Then the equations of motion can be written in Euler\u2013Poisson form as follows (Zuyev 2001): I1 _!1 \u00bc \u00f0I2 I3\u00de!2!3 \u00fe 1 \u00fe d1, \u00f01a\u00de I2 _!2 \u00bc \u00f0I3 I1\u00de!1!3 \u00fe 2 \u00fe d2, \u00f01b\u00de I3 _!3 \u00bc \u00f0I1 I2\u00de!1!2 \u00fe d3, \u00f01c\u00de _v1 \u00bc !3v2 !2v3, \u00f01d\u00de _v2 \u00bc !1v3 !3v1, \u00f01e\u00de _v3 \u00bc !2v1 !1v2, \u00f01f\u00de where !1, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001007_0020-7403(75)90014-4-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001007_0020-7403(75)90014-4-Figure5-1.png", + "caption": "FIG. 5. Trail (t).", + "texts": [ + " 3 and 4) shift respectively relative to the frame A and to the fork B, and the pitch angle fL changes, whenever there occurs an alteration either in the steering angle a or in the roll angle'\\. This rule has an exception as regards the pitch angle: since e (see equations (I)) can be made to vanish by choosing r, band 0 such that r sO - bcO = 0, it is possible to make fL equal to zero for all a and,\\. This fact can be restated in an interesting way by bringing a quantity called \"trail\" into the discussion, this quantity being defined as the distance t between the points T and U in Fig. 5, where T is the point of contact of the front wheel with the supporting plane when a = ,\\ = 0, and U is the point of inter section of this plane and of the axis of rotation of the fork B. Thus, (41) The vehicle is said to be in a steady turn when it is moving in such a way that the steering angle a, the roll angle ,\\ and the rear wheel rotation rate y have constant values, say a, ~ and~. Under these circumstances, it follows from equations (17), (33) and (34) that fL, Sand K also have constant values, say (1, Sand ~, with ~ = -~a(r+r' c~) d-l cO sec A" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000871_09544062jmes949-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000871_09544062jmes949-Figure8-1.png", + "caption": "Fig. 8 Illustrations of CS and FS teeth profiles", + "texts": [ + " Mechanical Engineering Science JMES949 \u00a9 IMechE 2009 by guest on January 20, 2015pic.sagepub.comDownloaded from of the section of the FS that contains the bolt pattern has been considered and has been computed through another FEM. The torsional stiffness of this portion of the FS is linear and is\u2248530 000Nm/rad. The teeth of the CS are assumed to be fully rigid because they have a much wider root tooth section than does the FS. The profile of this component is modelled as a rigid surface that restrains FS displacements. As shown in Fig. 8(a), the teeth profile of the FS has triangular shapes, whereas that of the CS is trapezoidal. The illustrative drawing shown in Fig. 8(b) demonstrates the meshing between the two components. From both illustrations, one can see that the root section of the CS teeth is much wider and \u2018robust\u2019 than that of the FS. Therefore, the flexional deformation of theCS teeth ismuch smaller than that of the FS. In order to replicate the complex meshing process between the FS and CS, two load cases are used. The first applies a prescribed displacement to the inner ring of theWG shown in Fig. 4 in order to deform the FS into its elliptical shape within the CS", + " The area comprised between the dashed curve (Kr = 38 000N/mm) and the dotted curve (Kr = 114 000N/mm) corresponds to the envelope of the torsional stiffness response of the FEM of the HD. As in the numerical model the balls are replaced by compression-only elements, the idea is to give to these elements a stiffness equivalent to the compressed balls. An average value for the stiffness of the balls was computed according to the relationship between force and displacement. For a low compression force of 100N, the displacement is about 2.6e-3mm, which gives Kr > 38 000N/mm. A high stiffness value was estimated to be three time this value, i.e. 114 000N/mm. Finally, results in Fig. 8 are presented for a low stiffness value, a high stiffness value, and a nearly infinite stiffness value. Figure 10 shows that the FEM can correctly reproduce the non-linear behaviour of a loaded HD. However, the torsional stiffness curve obtained from the FEM is slightly stiffer than the one provided by the manufacturer, especially for higher torques. Considering that the actual torsional stiffness of anHDcan vary by up to 15per cent comparedwith themanufacturer\u2019s specifications [15], the FEM of the HD is deemed accurate and efficient in providing rapid results" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001895_978-3-642-32448-2_11-Figure11.7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001895_978-3-642-32448-2_11-Figure11.7-1.png", + "caption": "Fig. 11.7 Cutting force calculation", + "texts": [ + " 11 Analysis of Industrial Robot Structure and Milling Process Interaction 253 Adding all disc levels together, a chart of the chip thickness h(\u03c6,z) is extracted, which forms the base of the force calculation. This is the most frequently performed arithmetic operation during the removal simulation. For the prediction of the cutting forces a standard cutting force model based on Altintas [3] is implemented, where the time delay responsible for the chatter is neglected. According to the chip discretization the cutter is sectioned into discs of height dz (cf. Fig. 11.7). In each disc e, Frta,j represents the force per tooth j in radial, tangential and axi- al direction, ( ), , \u00b7 \u00b7 , \u00b7jdz h z dz\u03d5= +rta j e c eF K K . 254 J. Bauer et al. Depending on the angular position of the tooth j of a disc, the corresponding chip thickness hj(\u03c6,z) is inserted. The cutting force coefficients Kc=[Krc, Ktc, Kac] and Ke=[Kre, Kte, Kae] need to be identified in advance. A transformation of Frta,j,e with T(\u03c6) and the subsequent summation over all teeth Nz and discs Ne results in the process force Fxyz,tool, given in a non-rotating tool coordinate frame, , , , 1 1 ( )\u00b7 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000059_j.conengprac.2008.05.004-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000059_j.conengprac.2008.05.004-Figure3-1.png", + "caption": "Fig. 3. Schematic of SLADe.", + "texts": [ + " The estimators make use of MEMS inertial sensors, a GPS receiver, a magnetometer and an ultrasonic altimeter to determine the required feedback signals. Simulation results confirm the effectiveness of the controller\u2013estimator combination and compare the results of the linear, decoupled estimators to those of a 16 state nonlinear kinematic state estimator (Hough, 2007). Finally, flight test results are presented to verify the practical performance of the controller. Consider the schematic overview of the aircraft in Fig. 3. The diagram shows the orientation of the defined body axis system with respect to the vehicle. The deflection angles of the eight flaps are denoted by the symbols d1\u2013d8. A positive flap deflection is defined as one that results in further blockage of the airflow in the lightly shaded regions of the left hand schematic of Fig. 3. The normalised angular velocities of the top and bottom rotors are denoted by the control variables d9 and d10, respectively. The symbols r, h, d, S and A represent the radius of the duct, the height of the centre of mass above the flap\u2019s centre of pressure, the radial distance between the centre of mass and the flap\u2019s centre of pressure, the area of a flap and the cross-sectional area of the duct, respectively. Nomenclature aL, aD flap lift and drag coefficients, respectively A area of the duct AD side-profile area CD side-profile drag coefficient d yaw moment arm from the flaps DX, DY axial and lateral coordinates of the drag force in body axes g gravitational force per unit mass h pitch and roll moment arms from the flaps Ix, Iy, Iz moments of inertia about the respective body axis system unit vectors kT, kM rotor thrust and moment coefficients, respectively K() feedback gain L() estimator gain m aircraft mass q dynamic pressure r radius of the duct S area of a flap T thrust force V duct airflow velocity magnitude L, M, N0 coordinates of the moment vector in body axes N, E, D coordinates of the position vector in the NED axis system P, Q, R coordinates of the angular velocity vector in body axes U, V, W coordinates of the translational velocity vector in body axes X, Y, Z coordinates of the force vector in body axes d1,y,d8 flap deflection angles d9, d10 top and bottom normalised rotor speeds dRF ; dPF ; dYF virtual roll, pitch and yaw actuators using flaps, respectively dTR; dYR virtual thrust and yaw actuator using rotors, respectively d0TR; dTR0 virtual thrust perturbation and trim value, respectively dY virtual yaw actuator for complementary control r air density t time constant F;Y;C Euler 3-2-1 angles of the body axis system with respect to the inertial space The aircraft is modelled as a six degree of freedom rigid body with aerodynamic, propulsion and gravitational forces and moments acting on it" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.14-1.png", + "caption": "FIGURE 6.14 A journal bearing.", + "texts": [ + " (1964) gave the following relationship for the critical Taylor number for an annulus with inner cylinder rotation and axial throughflow based on their experimental measurements, for a narrow gap annulus with Rez < 25. 2\u00f0a=b\u00de2\u00f0b\u2212a\u00de4 \u03a92 Ta2 \u00bc \u00f06:67\u00decr 2 21\u2212\u00f0a=b\u00de The term bearing typically refers to contacting surfaces through which a load is transmitted. Bearings may roll or slide or do both simultaneously. The range of bearing types available is extensive, although they can be broadly split into two categories: sliding bearings (see Figure 6.14), where the motion is facilitated by a thin layer or film of lubricant, and rolling element bearings, where the motion is aided by a combination of rolling motion and lubrication. Lubrication is often required in a bearing to reduce friction between surfaces and to remove heat. Here consideration will be limited to the fluid flow associated with one particular type of sliding bearings, rotating journal bearings. An introduction to rolling element bear ings is given in Childs (2004). The term sliding bearing refers to bearings where two surfaces move relative to each other without the benefit of rolling contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000726_icems.2009.5382754-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000726_icems.2009.5382754-Figure1-1.png", + "caption": "Fig. 1. Configuration of segment type SRM.", + "texts": [ + " Especially our segment type SRM has large torque ripples because the minimum torque is almost zero and the maximum torque is large. In this paper, we propose a torque ripple reduction method by increasing the number of phases. If we make the segment type SRM to poly-phase greater than three and excite two or more phases simultaneously, the torques produced by the two or more phases are superimposed in start and end of excitation, and the torque ripple is reduced. We analyze the effect of increasing the number of phases on the torque characteristics by 2-dimentional finite element method (FEM). Figure 1 shows construction of the novel segment type SRM. Segment cores are embedded in the aluminum rotor block and the stator has full pitch three phase windings. Features of this motor are as follows; (1) the average torque is larger, the radial force is smaller and so the vibration and acoustic noise are smaller than the VR type SRM because four poles among the six poles are always excited, (2) the iron loss is low because the magnetic path is short, (3) it has defective losses such as copper loss in the large coil end and eddy current loss in the aluminum rotor block" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002647_0954406212461326-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002647_0954406212461326-Figure3-1.png", + "caption": "Figure 3. The geometric relationship near crack boundary points.", + "texts": [], + "surrounding_texts": [ + "Based on the Hu\u2013Washizu variation integral principle,18 a variational equation of a structure under a different state is given as J\u00bc Z t2 t1 nZ v n ij\u00fePi _ i ui\u00fe ij e \"ij \"ij \u00fe ij 1 2 ui, j\u00feuj, i ij\u00fe _ui 1 2 i i i o dV \u00fe Z S1 gi gi uidS1\u00fe Z S2 ui ui\u00f0 \u00de gidS2 o dt\u00bc 0, \u00f01\u00de where is the body density, e is the strain energy density function, gi is the surface traction, V is the total volume of structure component, \"ij is the strain on the every section, ij is the stress on the every section, i is the velocity component, ui is the displacement component, Pi is the body force, S1 is the subjected boundary line of traction force, S2 is the subjected boundary line of displacement, and gi and ui are the traction force and displacement on the boundary line respectively. When a crack appears, the strain energy density usually increases for a beam dynamic system, the strain and stress also change, and the displacement also mutates. So it can be expressed by displacement maladjustment function f(x, z), if the body forces are inexistent then the boundary conditions are obtained as Displacement ux \u00bc z 1\u00fe f x, z\u00f0 \u00de\u00bd w x, t\u00f0 \u00de 0 , uy \u00bc 0, uz \u00bc 1\u00fe f x, z\u00f0 \u00de\u00bd w x, t\u00f0 \u00de Velocity vx \u00bc 0, vy \u00bc 0, vz \u00bc V x, t\u00f0 \u00de Strain \"xx \u00bc zS x, t\u00f0 \u00de, \"yy \u00bc \"zz \u00bc \"xx, \"xy \u00bc \"yz \u00bc \"xz \u00bc 0 Stress xx \u00bc zT x, t\u00f0 \u00de, xz \u00bc xz x, z, t\u00f0 \u00de, xy \u00bc zz \u00bc yy \u00bc yz \u00bc 0 Bodyforces Px \u00bc Py \u00bc Pz \u00bc 0, \u00f02\u00de where w(x, t) is the transverse vibration displacement, and S(x, t) and T(x, t) are the strain function and stress function related with crack respectively. Substituting equation (2) into equation (1), we can deduce that E I I2\u00f0 \u00de Q1 x\u00f0 \u00dew00 \u00feQ2 x\u00f0 \u00dew0 \u00feQ3 x\u00f0 \u00dew\u00f0 \u00de\u00bd 00 \u00fe R x\u00f0 \u00dew00 \u00bc 0, \u00f03\u00de where Q1(x)\u00bc I6/(I 2I2), Q2(x)\u00bc 2I5/(I 2I2), Q3(x)\u00bc I4/(I 2I2), R x\u00f0 \u00de \u00bc R A 1\u00fe f\u00f0 \u00dedA, I \u00bc R A z2dA, I2 \u00bcR A zdA, I4 \u00bc R A z2f 00 dA, I5 \u00bc R A z2f 0 dA, I6 \u00bc R A z2 1\u00fe f\u00f0 \u00de fdA, f 00 \u00bc @2f=@x2, E is the Young\u2019 modulus of elastomer. The displacement maladjustment function f(x, z) by deducing can be rewritten as: f x, z\u00f0 \u00de \u00bc 6 1 2 h2 \u00f0 \u00deL z L2 \u00fe h2=4\u00f0 \u00de ln x xcj j, \u00f04\u00de where \u00f0 \u00de \u00bc 0:6272 2 1:04533 3 \u00fe 4:5948 4 9:9736 5 \u00fe 20:2948 6 33:0351 7 \u00fe 47:1063 8 40:7556 9 \u00fe 19:6 10 is a polynomial of crack depth rate \u00bc a/h obtained by fitting experiment results. In equation (3), if the structure is a cracked beam that has rectangle section with the height h and width b, then I \u00bc bh3=12, I2 \u00bc 0,Q1 x\u00f0 \u00de \u00bc R x\u00f0 \u00de=A,Q2 x\u00f0 \u00de \u00bc 2R0 x\u00f0 \u00de=A,Q3 x\u00f0 \u00de \u00bc R00 x\u00f0 \u00de=A,A \u00bc bh: So the equation (3) can be expressed as C2 0 R x\u00f0 \u00dew\u00f0 \u00de \u00f04\u00de \u00fe R x\u00f0 \u00dew x, t\u00f0 \u00de 00 \u00bc 0, \u00f05\u00de where C2 0\u00bcEI/( A). Equation (5) is the vibration differential equation of the cracked beam. From Equation (5) it can be understood that this vibration response w(x, t) is only affected by R(x), in other words, R(x) affects w(x, t) directly. By considering the boundary and initial conditions, the solution of this differential equation (5) will be obtained. Dynamic response and dynamic characteristics of cracked gear The vibration differential equation (5) of successive beam with crack can be solved by using the separate variable method. By setting the vibration solution as w(x, t)\u00bcW(x)T(t) and putting it into equation (5), R x\u00f0 \u00deW x\u00f0 \u00de\u00bd \u00f04\u00de \u00fe\u00f0! n=C0\u00de 2 R x\u00f0 \u00deW x\u00f0 \u00de\u00bd \u00bc 0, \u20acT\u00f0t\u00de \u00fe \u00f0! n\u00de 2T\u00f0t\u00de \u00bc 0: \u00f06\u00de are obtained where, \u00f0! n=C0\u00de 2 \u00bc n 4 or ! n \u00bc C0 n 2 So the solution of the first equation in equation (6) is R x\u00f0 \u00deW x\u00f0 \u00de \u00bc An cos nx \u00fe Bn sin nx \u00fe Cn ch nx \u00feDnsh nx : at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from The solution of the second equation in equation (6) is T t\u00f0 \u00de \u00bc En0 cos ! nt \u00fe Fn0 sin ! nt : So the complete solution of the successive body is w x,t\u00f0 \u00de \u00bcW x\u00f0 \u00deT t\u00f0 \u00de \u00bc X 1=R x\u00f0 \u00de:\u00bdAn cos\u00f0 nx\u00de\u00feBn sin\u00f0 nx\u00de \u00feCn ch\u00f0 nx\u00de\u00feDn sh\u00f0 nx\u00de \u00bdGn0 sin\u00f0! nt\u00fe \u00de \u00f07\u00de Where Gn0 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi E2 n0 \u00fe F2 n0 q are constant. Every constant (An, Bn, Cn, Dn,, En0, Fn0) can be gained according to the boundary supported conditions and initial conditions, that is, the special solution response of equation (5) can be obtained. The dynamic characteristics of gear body are solved by setting gear tooth as cantilever cracked beam, set \u2019(x)\u00bc 1/R(x), then the inherence function (vibration shape) W(x) can be expressed as W x\u00f0 \u00de \u00bc \u2019 x\u00f0 \u00de An cos nx \u00fe Bn sin nx \u00fe Cn ch nx \u00feDn sh nx : \u00f08\u00de The conditions of restriction in cantileverbeam are: Fixed-end: displacement W(x) j x\u00bc 0\u00bc 0, rotational angle @W x\u00f0 \u00de=@xjx\u00bc0 \u00bc 0: \u00f09\u00de Free-end: bend-moment @2W(x)/@x2jx\u00bcL\u00bc 0, shearing force @3W x\u00f0 \u00de=@x3jx\u00bcL \u00bc 0: \u00f010\u00de If equation (8) is substituted into equation (9), then, An \u00fe Cn \u00bc 0 An\u2019 0 0\u00f0 \u00de \u00fe Bn n\u2019 0\u00f0 \u00de \u00fe Cn\u2019 0 0\u00f0 \u00de \u00feDn n\u2019 0\u00f0 \u00de \u00bc 0 \u00f011\u00de If equation (8) is substituted into equation (10), then, Anb31 \u00fe Bnb32 \u00fe Cnb33 \u00feDnb34 \u00bc 0 Anb41 \u00fe Bnb42 \u00fe Cnb43 \u00feDnb44 \u00bc 0 \u00f012\u00de where b31\u00bc \u201900 L\u00f0 \u00de \u2019 L\u00f0 \u00de n 2h i cos nL 2\u20190 L\u00f0 \u00de nsin nL b32\u00bc \u201900 L\u00f0 \u00de \u2019 L\u00f0 \u00de n 2h i sin nL \u00fe2\u20190 L\u00f0 \u00de n cos nL b33\u00bc \u201900 L\u00f0 \u00de\u00fe\u2019 L\u00f0 \u00de n 2h i ch nL \u00fe2\u20190 L\u00f0 \u00de n sh nL b34\u00bc \u201900 L\u00f0 \u00de\u00fe\u2019 L\u00f0 \u00de n 2h i sh nL \u00fe2\u20190 L\u00f0 \u00de n ch nL b41\u00bc \u201900 0 L\u00f0 \u00de 3\u20190 L\u00f0 \u00de n 2h i cos nL 3\u201900 L\u00f0 \u00de \u2019 L\u00f0 \u00de n 2h i n sin nL b42\u00bc \u201900 0 L\u00f0 \u00de 3\u20190 L\u00f0 \u00de n 2h i sin nL \u00fe 3\u201900 L\u00f0 \u00de \u2019 L\u00f0 \u00de n 2h i n cos nL b43\u00bc \u2019000 L\u00f0 \u00de\u00fe3\u20190 L\u00f0 \u00de n 2h i ch nL \u00fe 3\u201900 L\u00f0 \u00de\u00fe\u2019 L\u00f0 \u00de n 2h i n sh nL b44\u00bc \u201900 0 L\u00f0 \u00de\u00fe3\u20190 L\u00f0 \u00de n 2h i sh nL \u00fe 3\u201900 L\u00f0 \u00de\u00fe\u2019 L\u00f0 \u00de n 2h i n ch nL Combining equations (11) and (12), Bj jfAg \u00bc 0 \u00f013\u00de is obtained where: {A}\u00bc {An Bn Cn Dng T. So the eigenequation is jBj \u00bc 0, \u00f014\u00de where Bj j \u00bc 1 0 1 0 \u20190 0\u00f0 \u00de n\u2019 0\u00f0 \u00de \u20190 0\u00f0 \u00de n\u2019 0\u00f0 \u00de b31 b32 b33 b34 b41 b42 b43 b44 , The eigenvalue (eigenroot) is obtained by solving equation (14), and the natural frequency ! n\u00bcC0 ( n) 2 also is obtained for the cracked gear. By putting n into equation (13), the constant ratio is obtained, then, putting it into equation (8), the natural function of successive body are obtained, i.e. vibration shape. Theoretical basis of structure fracture analysis Crack modes and stress intensity factor Usually the propagation ways of crack under the external force can be divided into three forms: (a) opening model (I): the crack is opened and extend under the tensile stress which is perpendicular to the crack surface; (b) sliding model (II): the crack is slided and it extends under the shear stress which is parallel to the crack surface and perpendicular to the crack at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from tip; (c) tearing model (III): the crack is torn and extends under the shear stress which is both parallel to the crack surface and parallel to the crack front. Three different crack propagation modes are shown as Figure 1. In actual engineering, for involute gear, the 3D crack of tooth root is composite with I, II and III modes. \u2018\u2018Intensity\u2019\u2019 of stress field near the crack tip can be expressed as a factor that depends on crack geometry and load conditions, for different crack modes, they are denoted as KI, KII, KIII respectively, so the stress distribution in crack tip can usually be expressed as: ij\u00f0r1, 1\u00de\u00bc 1ffiffiffiffiffiffiffiffiffi 2 r1 p KIf I ij\u00f0 1\u00de\u00feKIIf II ij \u00f0 1\u00de\u00feKIIIf III ij \u00f0 1\u00de n o \u00f015\u00de here KI, KII, KIII are called stress intensity factor (SIF), they stand for the intensity of stress field and are related to crack, geometry of crack body (including size and shape) and external load conditions. fIij\u00f0 1\u00de, f II ij \u00f0 1\u00de, f III ij \u00f0 1\u00de stand for the geometrical relationship of KI, KII and KIII modes respectively in the crack tip polar coordinate of crack. The definition of 2D crack stress intensity factor generally is: KI KII KIII 2 4 3 5 \u00bc lim r1!0 ffiffiffiffiffiffiffiffiffi 2 r1 p yy\u00f0r1, 0\u00de xy\u00f0r1, 0\u00de yz\u00f0r1, 0\u00de 2 4 3 5 \u00f016\u00de The correlative parameters of equations (15), (16) are shown in Figures 2 and 3. For 3D crack problems, the gradual stress field in crack boundary can be expressed as follows when it is accurate to 1= ffiffi r p n \u00bc KIffiffiffiffiffiffiffi 2 r p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00fe 1 2 cos r 1 1\u00fe \u00f01 \u00de\u00f02\u00fe \u00de 2 2 KIIffiffiffiffiffiffiffi 2 r p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 2 cos r 1 2\u00fe \u00f01\u00fe \u00de\u00f02 \u00de 2 2 z \u00bc KIffiffiffiffiffiffiffi 2 r p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00fe 1 2 cos r 1 1 \u00f01 \u00de\u00f02\u00fe \u00de 2 2 \u00fe KIIffiffiffiffiffiffiffi 2 r p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 2 cos r \u00f01\u00fe \u00de\u00f02 \u00de 2 3 t \u00bc 2 ffiffiffiffiffiffiffi 2 r p 1ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 cos p KI ffiffiffiffiffiffiffiffiffiffi \u00fe 1 p KII ffiffiffiffiffiffiffiffiffiffi 1 ph i nz \u00bc KIffiffiffiffiffiffiffi 2 r p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 2 cos r \u00f01\u00fe \u00de\u00f02 \u00de 2 3 \u00fe KIIffiffiffiffiffiffiffi 2 r p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00fe 1 2 cos r 1 1\u00fe \u00f01 \u00de\u00f02\u00fe \u00de 2 2 nt \u00bc KIIIffiffiffiffiffiffiffi 2 r p 1 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 2 cos r tz \u00bc KIIIffiffiffiffiffiffiffi 2 r p 1 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00fe 1 2 cos r \u00f017\u00de KI \u00bc lim r!0 ffiffiffiffiffiffiffiffiffiffi 2 r p zj \u00bc0,KII \u00bc lim r!0 ffiffiffiffiffiffiffiffiffiffi 2 r p nzj \u00bc0, KIII \u00bc lim r!0 ffiffiffiffiffiffiffiffiffiffi 2 r p tzj \u00bc0: at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from When the crack surface is loaded under uniform normal stress, KI \u00bc p0 E\u00f0k\u00de ffiffiffiffiffi b a r ffiffiffiffiffiffi , p KII \u00bc KIII \u00bc 0 When the crack surface is loaded under uniform shearing stress, KI\u00bc0 KII\u00bck2k 02q0 ffiffiffiffiffiffiffi = p a sin sin \u2019 k02K\u00f0k\u00de \u00f0k2\u00fe k02\u00deE\u00f0k\u00de \u00fe b cos cos\u2019 \u00f0 k2\u00deE\u00f0k\u00de k02K\u00f0k\u00de KIII\u00bck2k 02\u00f01 \u00deq0 ffiffiffiffiffiffiffi = p a cos sin \u2019 \u00f0k2 \u00deE\u00f0k\u00de\u00fe k02K\u00f0k\u00de b sin cos\u2019 \u00f0k2\u00fe k02\u00deE\u00f0k\u00de k02K\u00f0k\u00de \u00f018\u00de If point P and point P0 are both located in normal plane, so \u00bc 0, \u00bc 1, \u00bc sec( ), then, equation (17) will be degenerate to the formula that is same with the situation of plane strain in normal plane. stands for the angle between the shearing stress on crack surface and the abscissa axis. Boundary element method of stress intensity factor The SIF\u2019s solution of crack front is a complicated process in structures, especially 3D crack. So far, in various studies, the relative displacement method and J integral method are widely used to analyze the numerical simulation of SIF. Using the relative displacement method to solve SIF, the displacement of every point at crack front only need to be solved, so the method is fit for solving SIF in boundary element analysis. The analysis process of boundary element method(BEM) is as follows: Ignoring the body forces and conducting static analysis of 3D elastomer, the boundary equation can be gained as follows: Cij\u00f0P\u00deUj \u00f0P\u00de \u00fe Z s Uj \u00f0Q\u00deT ij\u00f0P,Q\u00dedS\u00f0Q\u00de \u00bc Z s Tj\u00f0Q\u00deU ij\u00f0P,Q\u00dedS\u00f0Q\u00de \u00f019\u00de where Uj is displacement; Tj is surface force; U ij is basic solution of displacement; T ij is basic solution of surface force. If the boundary is divided into m elements, and the number of element nodes is n, so element geometry, displacement and surface force interpolation function can be defined as: Xi\u00f0 , \u00de \u00bc Xn k\u00bc1 Nk\u00f0 , \u00deXik Ui\u00f0 , \u00de \u00bc Xn k\u00bc1 Nk\u00f0 , \u00deUik Ti\u00f0 , \u00de \u00bc Xn k\u00bc1 Nk\u00f0 , \u00deTik 8>>>>>>>>>< >>>>>>>>>: \u00f020\u00de where Xik is the coordinate value of the Kth node along i direction, Uik is the displacement value of the Kth node along i direction, Tik is the surface force value of the Kth node along i direction, interpolation function can be defined by natural coordinates. Substituting equation (20) into (19), the boundary integral discrete equations are established as follows for every boundary point P: Cij\u00f0P\u00deUj \u00f0P\u00de \u00fe Xm L\u00bc1 Xn k\u00bc1 Z 1 1 Z 1 1 Nk\u00f0 , \u00deT ij\u00f0P,Q\u00de ( Jj jd dn U \u00f0k\u00de i o \u00bc Xm L\u00bc1 Xn k\u00bc1 Z 1 1 Z 1 1 Nk\u00f0 , \u00deU ij\u00f0P,Q\u00de Jj jd dn T \u00f0k\u00de i ( ) \u00f021\u00de where J is the Jacobi inequality. The matrix expression of the equations (21) is: A\u00bd Uf g \u00bc B\u00bd Tf g \u00f022\u00de According to the known boundary conditions, the unknown displacement and surface force will be obtained by solving above equations. After solving the boundary displacement, the magnitude of SIF can be gained. at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from Simulation of dynamic characteristics for cracked gear The gear tooth as cantilever beam is shown in Figure 4, where the tooth length is L, the crack location is xc, the crack depth is a, and the gear width is b. This outline of tooth is involute and different from the beam with rectangle section, where the thickness of gear root is h1 and the thickness of the top is h2. Here the simulated gear is modulus m\u00bc 10 and tooth number z\u00bc 80. Based on the meshing theory of gear, the gear parameters can be gained as h1\u00bc 20.3mm, h2\u00bc 10.7mm, L\u00bc 28.7mm, and b\u00bc 15mm. From the above analyses, it is possible to obtain the solution of eigenroot n, natural frequency and vibration shape, and the influences of crack on both natural frequency and vibration shape are analyzed. The influence of crack on natural frequency The Figures 5, 6, and 7 show the changes of first ninthorders natural frequency along with the crack depth when the crack locations are at 0.25L, 0.5L and 0.75L, respectively. From these figures it can be seen that these natural frequencies will drop with increase of the crack depth, and the deeper the crack is, the more serious the drops of natural frequency are. It is also possible to see that the drops of lower order natural frequency are obvious, such as the first fifth-orders, and the drops of higher order natural frequency are slow with increase of the crack depth. Besides, the drop extent of natural frequencies is affected more by the location of crack, such as the first-order natural frequency, and when the crack locates at near the fixed-end (Figure 5, xc\u00bc 0.25L) its drop extent is bigger, while near the free-end (Figure 6 and 7) its drop extent is smaller. In a word, the influence of crack location on the lower natural frequency (such as the first fifth-orders) is bigger, but it is smaller on the higher one, and the drop extent is bigger when the crack near the fixed-end while smaller when near the free-end. All these are agreement to the actual situation in engineering. In fact the stiffness will be declined when crack appears, so the natural frequency of gear will show itself drop. To the cantilever-beam, the whole system is supported by using the fixed-end, the damage of support stiffness is big when the crack is near the at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from fixed-end and it is small when the crack is near the freeend, and it even is no effect when the crack is at the end of the free-end. The influence of crack on vibration shape Figure 8 shows the first ninth-orders vibration shape of gear tooth when the crack locates at 0.25L, the rate of crack depth is 0 and 0.3, respectively (the rate of smooth curve is \u00bc 0, and the rate of zigzag one is \u00bc 0.3). From Figure 8 it can be seen that the vibration shape of gear without crack is a smooth curve, and when the number of order is higher than two, the curves show equal shape of vibration. While the vibration shape will change and it will mutate at the crack when a crack appears. Because of the effect of crack, at NORTH CAROLINA STATE UNIV on October 17, 2014pic.sagepub.comDownloaded from the vibration shape of the place without crack will decline sharply (caused by the crack). All of these are valuable to identify the location of crack. Figure 9 shows the first ninth-orders vibration shape of gear tooth when the crack locates at 0.50L, the rate of crack depth is 0 and 0.6, respectively (the rate of smooth curve is \u00bc 0, and the rate of zigzag one is \u00bc 0.6). From Figure 9 it is possible to see that the vibration shape changes sharply when the crack appears, especially the big crack, the vibration shapes show a sawtooth at the position of crack. From a macroscopic perspective, the vibration phase of the crack in the fixed-end is inverse to that without crack, i.e. the phase discrepancy is 180 . The shape amplitude declines near the crack and becomes unequal, while the node number of vibration shape is reduced by one because of crack, which is a difference between the vibration shape of big crack and that of small one. Besides, the mutation of the odd number order of vibration shape in crack (center of beam) is bigger than that of the even number order. The reason is that the crack is just at the node of vibration shape. Therefore, in order to identify the location of crack, attention should be paid not only to the mutation of the vibration shape but also to the place of vibration shape node. The finite element simulation of dynamic characteristics for cracked gears" + ] + }, + { + "image_filename": "designv11_3_0003469_j.oceaneng.2013.01.001-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003469_j.oceaneng.2013.01.001-Figure15-1.png", + "caption": "Fig. 15. Simplified model of 6-legged robot.", + "texts": [ + " This fact means that the 4-legged robot cannot move forward (the positive X axis direction) and drift with the tidal current. From this result, we assure that the proposed method can be utilized to analyze the tidal current endurance performance of an underwater walking robot. Finally, to prove the application capability of the proposed method, we have performed the simulation for mobility and agility analysis of a 6-legged robot in underwater environment. For this simulation, the simplified model of 6-legged robot of Fig. 15 (the model of Fig. 7\u00fetwo legs) has been used and the underwater environment has been defined as the same parameter In addition to the physical intuitions of the example case I, II, and III, we can easily expect several results: (1) all acceleration bounds are bigger than ones of the 4-legged robot, (2) the tidal current endurance capability is larger than the example case III. Figs. 16, 17, and Table 8 show the mobility and agility of the 6- legged robot in ground and underwater environment. Here we know that the maximum acceleration along X axis in the 2 knots case has 4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000178_s12008-007-0003-7-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000178_s12008-007-0003-7-Figure2-1.png", + "caption": "Fig. 2 Studied assemblies: a ring screwed to support, b support screwed to ring using cylindrical casing", + "texts": [ + " The bearing used features 72 screws uniformly distributed on the rings. Each ring has 36 screws (M16, grade HR 10.9), 6 of which are equipped with uniaxial strain gauges. The equipped screws were located according to Fig. 1b. That allows evaluating the load distribution in the joint. Symmetry of the joint loading may also be checked. 2.2 Studied assemblies Several configurations were tested. The choice of configurations resulted from a survey among manufacturers and users of slewing bearings. Figure 2 sums up the most typical configurations. The first assembly features tapped mountings. The second one uses tapped rings. Here the screws tighten the mounting plate against the ring. One drawback of that assembly lies in the fact that the screws are short, thus the cyclic loading is higher. The most often used and most economical remedy lies in adding spacers (Fig. 2b). Screw rigidity is reduced. This lessens possible stiff points and makes load distribution more uniform. From here, we only deal with the two assemblies of Fig. 2. Presented hereafter are the numerical models which were used to better understand the behavior of the joints. The first step consists in validating the finite element models using experimental results. Secondly, geometrical and physical parameters are changed in order to analyse the assembly\u2019s behavior sensitivity to them. 3.1 Three-dimensional FE model Geometrical symmetry of the assembly allows reducing the modeled volume to the most loaded angular sector, although the loading does not comply to the symmetry" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002171_2041301710394920-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002171_2041301710394920-Figure3-1.png", + "caption": "Fig. 3 Illustration of wheel/rail", + "texts": [ + " Using a given \u2018mother\u2019 profile of wheel with basic standard parameters such as height and thickness of flange, the design can be carried out to optimize this profile. This is the basis of Proc. IMechE Vol. 225 Part F: J. Rail and Rapid Transit at SYRACUSE UNIV LIBRARY on November 23, 2013pif.sagepub.comDownloaded from the design methodology which will now be described. The following assumptions are made. 1. The wheel and rail are both rigid bodies. 2. The rail profile must be continuous and convex. 3. Symmetry of left and right wheels on symmetric rails, assumed only for profile design. As shown in the Fig. 3, a wheelset contacts a pair of rails with a given profile. An axis system is defined as follows: lateral axis parallel to the common tangent of the left and right rail crowns, positive towards the right. The vertical axis z is positive upwards. The co-ordinate values of left and right wheel profile, and left and right rail profile are defined as (Ywl, Zwl) and (Ywr, Zwr), and (Yrl, Zrl) and (YrrZrl), respectively. Figure 4 shows the modified (or optimized) RRDF. Ys is the lateral shift of wheelset, and R is the half of left and right the rolling radius difference between left and right wheel. R = D(Ys) is defined as a target function of the designed profile. For a movement of Ys, the wheel treads contact rails at points indicated in Fig. 3 by \u2018I \u2019 on both sides (as shown in Fig. 3) with a roll angle of the wheelset \u03c6(Ys). The gradients on the contact points are Kl(Ys) and Kr(Ys), respectively, for left and right. The co-ordinate values of the contact points I are (Ywl(Ys), Zwl(Ys)) and (Ywr(Ys), Zwr(Ys)) for the left and right wheel profile, respectively, and (Yrl(Ys), Zrl(Ys)) and (Yrr(Ys), Zrr(Ys)) for the left and right rail profile, respectively. Now the wheelset is further moved with a step h. The lateral movement of wheelset becomes Ys + h, and the rolling radius differences becomes D(Ys + h), and the roll angle becomes \u03c6(Ys + h). The contact points moves from I to II (as shown in Fig. 3). The co-ordinate values of the contact points \u2018I \u2019 are (Ywl(Ys + h), Zwl(Ys + h)) and (Ywr(Ys + h), Zwr(Ys + h)) for the left and right wheel profile, respectively, and (Yrl(Ys + h), Zrl(Ys + h)) and (Yrr(Ys + h), Zrr(Ys + h)) for the left and right rail profile, respectively. The roll angles can be obtained according to the geometric relationship as follows \u03c6(Ys) = arctg ( Zwr(Ys) \u2212 Zrr(Ys) \u2212 (Zwl(Ys) \u2212 Zrl(Ys)) Yrr(Ys) \u2212 Yrl(Ys) ) \u03c6(Ys + h) = arctg \u239b \u239c\u239c\u239c\u239d Zwr(Ys + h) \u2212 Zrr(Ys + h) \u2212(Zwl(Ys + h) \u2212 Zrl(Ys + h)) Yrr(Ys + h) \u2212 Yrl(Ys + h) \u239e \u239f\u239f\u239f\u23a0 (1) Proc. IMechE Vol. 225 Part F: J. Rail and Rapid Transit at SYRACUSE UNIV LIBRARY on November 23, 2013pif.sagepub.comDownloaded from As the contact point, the gradients on wheel and rail profiles are related as follows arctg(Z\u0307rl(Yrl(Ys)) = arctg(kl(Ys)) \u2212 \u03c6(Ys) (2) arctg(Z\u0307rr(Yrr(Ys)) = arctg(kr(Ys)) + \u03c6(Ys) (3) where \u2018rl\u2019 and \u2018rr\u2019 represent the left rail and right rail, respectively. According to the geometric relationship in Fig. 3, one has Ywl(Ys + h) = (Yrl(Ys + h) \u2212 (Ys + h))\u2217 cos(\u03c6(Ys + h)) \u2212 zwl(Ys + h)\u2217 sin(\u03c6(Ys + h)) (4) Ywr(Ys + h) = (Yrr(Ys + h) \u2212 (Ys + h))\u2217 cos(\u03c6(Ys + h)) \u2212 zwr(Ys + h)\u2217 sin(\u03c6(Ys + h)) (5) where \u2018wl\u2019 and \u2018wr\u2019 represent the left- and right-wheel profile, respectively. h is a deviation in calculation. If h \u2192 0, the gradient on\u2018I\u2019 and\u2018II\u2019 become the same, the section \u2018I\u2013II\u2019 on the wheel profile can be treated as a tangent line with constant gradient. The following is obtained kl(Ys) = Zwl(Ys + h) \u2212 Zwl(Ys) Ywl(Ys + h) \u2212 Ywl(Ys) (6) kr(Ys) = Zwr(Ys + h) \u2212 Zwr(Ys) Ywr(Ys + h) \u2212 Ywr(Ys) (7) The increase of D(Ys) can be obtained by D(Ys + h) \u2212 D(Ys) = Zwl(Ys + h) \u2212 Zwl(Ys) \u2212 (Zwr(Ys + h) \u2212 Zwr(Ys)) (8) There are 15 unknowns in the above eight equations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002273_1350650111403580-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002273_1350650111403580-Figure2-1.png", + "caption": "Fig. 2 Non-circular three-lobe bearing geometry", + "texts": [ + " The bearing configurations undertaken for this study are two-, three- and four-lobe finite bearings of symmetric geometries shown in Fig. 1. A non-circular lobed bearing is essentially an assembly of two or more partial circular arc bearings referred to as the lobes. The performance characteristics of the each partial bearing are then assembled, to obtain the overall bearing characteristics. A schematic diagram of a self-acting three-lobe journal bearing configuration and geometric space between journal and ith lobe, along with the coordinate systems used in the analysis, are shown in Fig. 2. Proc. IMechE Vol. 225 Part J: J. Engineering Tribology at UNIV OF PITTSBURGH on June 21, 2015pij.sagepub.comDownloaded from The modified Reynolds equation under micropolar lubrication, with the usual lubrication assumptions for an incompressible fluid flow, is the starting point for analysis. The modified form of the Reynolds equation for micropolar fluids can be expressed as [24] @ @x N , , h\u00f0 \u00de @P @x \u00fe @ @z N , , h\u00f0 \u00de @P @z \u00bc 6U @h @x \u00f01\u00de where N , , h\u00f0 \u00de \u00bc h3 \u00fe 12 2h 6N h2 coth Nh 2 \u00f02\u00de As tends to zero, equation (2) reduces to N , , h\u00f0 \u00de \u00bc h3 \u00f03\u00de Thus, equation (1) becomes @ @x h3 @P @x \u00fe @ @z h3 @P @z \u00bc 6U @h @x \u00f04\u00de which is the classical Reynolds equation for steadystate Newtonian fluid [40]", + " At higher values of lm, the effect of substructures become less and less significant and in the limiting case when lmapproaches infinity, the effect of individuality of the substructure is lost and the lubricant reduces to a Newtonian one. In this study, the results are for bearing eccentricity ratio (\") equal to 0.5 and aspect ratio ( ) of 1.0. The effect of lubricant supply grooves has been disregarded [41]. It has been assumed that the leading boundary i 1of any lobe begins from its junction with the preceding lobe, as shown in Fig. 2. To check the accuracy of the developed numerical model, the variation of load capacity as a function of eccentricity ratio obtained for a circular bearing lubricated with Newtonian as well as micropolar *This study, without groove. yMalik et al. [41] data, without groove. Sommerfeld number is modified according to the non-dimensional parameters of this study. Proc. IMechE Vol. 225 Part J: J. Engineering Tribology at UNIV OF PITTSBURGH on June 21, 2015pij.sagepub.comDownloaded from lubricants are compared with the published results [26] in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.3-1.png", + "caption": "FIGURE 6.3 Boundary layer flow over a rotating cylinder with superposed axial flow.", + "texts": [ + " The physical reason for this dependency is due to processes in the boundary layer where the fluid due to the no-slip condition co-rotates in the immediate vicinity of the wall and is therefore subject to the influence of strong \u201ccentrifugal forces.\u201d As a result, the processes of separation and transition from laminar to turbulent flow are affected by these forces and therefore drag too. The boundary layer on a rotating body of revolution in an axial flow consists of the axial component of velocity and the circumferential component due to the no-slip condition at the body surface. The result is a skewed boundary layer as illustrated in Figure 6.3 for the case of a cylindrical body. The thickness of the boundary layer increases as a function of the rotation parameter, \u03bbm, which is given by \u03a9a \u03bbm \u00bc \u00f06:1\u00de uz;\u221e where \u03a9 is the angular velocity of the cylinder, a is the radius of the cylinder, and uz,\u221e is the free-stream axial velocity. Steinheuer (1965) found the following relationship between the axial and rotational velocity for the boundary on a rotating cylinder with axial flow u uz\u00bc 1\u2212 \u00f06:2\u00de \u03a9a uz;\u221e A common requirement in practical applications is the need to quantify the power required to overcome the frictional drag of a rotating shaft" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001365_s11370-010-0081-4-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001365_s11370-010-0081-4-Figure6-1.png", + "caption": "Fig. 6 Cubic splines and circle segments are used to generate more complex trajectories for the rover", + "texts": [ + " The rover communication interface provides a central interface program to the higher levels of the system control. Rover motion planning The level A controller of the rover provides different types to control it\u2019s motion, i.e. velocity control (left and right wheels control, jog-rate etc.), control of certain distances at defined velocities, as well as control of more complex trajectories which can be assembled from simpler trajectory parts. Possible parts constituting a trajectory can be segments of circles as well as cubic splines which reach a certain point with given heading direction (Fig. 6). The general concept of motion planning of the rover consists of three main steps: \u2013 Depending on the task, a trajectory is planned either automatically as it is the case during the automatic docking to the lander mockup, or controlled by the operator from ground. \u2013 The trajectory is transferred to the controller which commands the motion system and the specific velocities for the left and right wheel system (differential drive). \u2013 During the motion execution, the motion is monitored via the odometry system which estimates a rough information of the position and heading direction of the rover" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003469_j.oceaneng.2013.01.001-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003469_j.oceaneng.2013.01.001-Figure3-1.png", + "caption": "Fig. 3. Vector diagram and link coordinates of underwater leg.", + "texts": [ + " The dynamic equation of k-th leg with n degree of freedom is described as sk \u00bcMk\u00f0qk\u00de \u20acqk\u00feHk\u00f0qk, _qk\u00de\u00feGLk\u00f0qk\u00de\u00fe JT k \u00f0qk\u00deFk\u00feDLk\u00f0qk, _qk\u00de \u00f04\u00de where tkARn is the joint torque vector, MkARn n is the inertia matrix including added mass, \u20acqkARn is the joint acceleration vector, HkARn is the Coriolis and centrifugal term caused by rigid body and added mass, GLkARn is the buoyancy and gravity term, JkAR6 n is the Jacobian matrix, FkAR6 is the force and moment exerted on the seabed by leg\u2019s end-effector, and DLkARn is the hydrodynamic drag and lift torque without shear drag term. In this study, these hydrodynamic forces are described by the existing study (Jun et al., 2006) utilizing the link coordinate system shown in Fig. 3. Through the study, the drag force acting on a thin slice of the j-th link with small thickness of dl can be approximated as a vector equation, which is represented in i-th frame by difDj \u00bc 1 2 rCDjdpj: ivn j : ivn j dl, \u00f05\u00de where difDj is the drag force acting on a slice in j-th link expressed in i-th coordinate system and CDj is 2-dimensional drag coefficient of j-th link. The dpj is the projected length of slice to the vector normal to ivn j . The ivn j is the translational velocity component of slice normal to the longitudinal axis of j-th link" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001855_j.compstruc.2008.12.014-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001855_j.compstruc.2008.12.014-Figure12-1.png", + "caption": "Fig. 12. Spatial bending of a circular rod.", + "texts": [ + " The results of the linear model for the X component of the displacement exhibit a fair accuracy only for low loads (|F R2/E I| 1), while for higher loads the inaccuracy increases. The nonlinear model agrees very well with the analytical results, for both compression and tension, even for extremely high deformations. Fig. 11 presents the shape of the rod after deformation, for two positive (tension) load magnitudes and two negative (compression) ones. For reference, the undeformed rod axis is also shown (zero loading). The numerical results of the present model exhibit good agreement with the analytical results, even for extremely large deformations. Fig. 12 presents the geometry of an undeformed cantilevered circular rod, lying in the plane X\u2013Z. The rod is clamped at its root and a concentrated force F, normal to the plane of the rod (in the Y direction), is applied at the free tip. The rod radius is R and its opening angle is equal to 45 . It has a square cross-section with bending stiffness (E I) about its principle axes. This problem was analyzed by Bathe and Bolourchi [15] using a finite element approach. Fig. 13 presents the components of the tip displacement in the directions (X,Y,Z), as functions of the normalized load, (F R2/E I)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002362_tmag.2012.2199094-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002362_tmag.2012.2199094-Figure5-1.png", + "caption": "Fig. 5. Slotted stator for rotation with slotless winding for translation in the airgap.", + "texts": [ + " However, the ratio between the flux seen by a coil in the circumferential direction and a coil in the axial direction can be altered by adjusting the ratio between the height of the axially magnetized PM and the pole pitch, . B. Stator Configuration for 2-DoF Actuator The stationary part of the actuator consists of two orthogonal winding sets, one oriented in the circumferential- and one in the axial-direction. To limit the complexity of the stator, a combination of one slotless, and one slotted winding is selected as illustrated in Fig. 5. The slotless winding for translation is inserted in the airgap of the slotted stator with the winding for rotation. The teeth and tooth-tips provide a high-permeable flux path in the axial direction and can be regarded as the back-iron of the slotless winding. As the stator is slotted in one direction, cogging force/torque is only present in one direction of motion. As the magnetic airgap is large due to the inserted winding for translation, the cogging is expected to be low. Due to the structure and the two degrees of freedom provided by the actuator, 3D models are inevitable", + " However, as the windings are orthogonal and oriented in the and the direction, 2D models in the axisymmetric and polar coordinate system can be exploited to approximate the magnetic fields. Therefore, two semi-analytical models are derived based on Fourier analysis as presented in [10] to calculate the translational and rotational performance. A. 2D Model in Axisymmetric Coordinate System A 2D model in the axisymmetric coordinate system is derived to approximate the translational performance of the actuator. The model, as illustrated in Fig. 6(a), is a simplification of the structure as shown in Fig. 5. Due to the selected coordinate system, the model is invariant in the circumferential direction. As such, the slots in the axial direction are not taken into account, hence, the stator back-iron is modeled as a soft-magnetic cylinder with infinite permeability. Three cylindrical concentric regions are defined as shown in Fig. 6: I The non-magnetic shaft of the mover. II The PM-array of the mover. III The airgap and slotless winding for translation. To calculate the magnetic fields due to the PM-array and the coils, the magnetic vector potential is exploited" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000620_1.3453275-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000620_1.3453275-Figure1-1.png", + "caption": "Fig. 1 3GO-degree bearing operating under starved conditions (a), equivalent partial bearing (b)", + "texts": [ + "/s) V = peripheral velocity of shaft, m/s (ft/ min) Journal of lubrication Technology W t = load perpendicular to line centers, N (lbf ) W = dimensionless load = (W/Il U L)/(C/ 1';)2 Wr = dimensionless load along line of centers = (Wr /Il UL )/C/r;)2 Wt = dimensionless load perpendicular to line of centers = (Wtlll U L)/(C/I';)2 Z = dimensionless distance along bearing = z/(L/2) e = eccentricity, m (in.) f = coefficient of friction h = oil film thickness = C [1 + \u20ac cos ('Y + 0)], m (in.) k = permeability, m2 (in.2) p = pressure in porous matrix, Pa (psi) p* = pressure in oil film Pa (psi) qb = oil flow from ends of porous matrix, m3/s (in3/s) qo == oil flow from clearance gap, m3/s (in.3/ s) 1', 0, z = cylindrical coordinates with origin at center of bearing, Fig. 1 s = slip parameter = q,1/2/ c< u = x-component of the oil velocity in the clearance gap, m/s (in./s) w = z-component of the oil seepage velocity in the porous matrix, m/s .(in./8) x, y, z = rectangular coordinate with origin at center of bearing, Fig. 1 q, = permeability parameter, k/C2 c< = dimensionless slip coefficient {3 = active film arc, degrees 'Y = angle from line of centers to beginning of oil film measured in the direction of shaft rotation, degrees Dnj '\" Kronecker delta \u20ac '\" eccentricity ratio, e/C 11 = absolute viscosity, Pa-s (lbf-s/in.2) Th '\" shear stress on surface of shaft, Pa (psi) '\" = attitude angle, degrees JANUARY 1979, VOL 101 / 39 Downloaded From: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jotre9/28621/ on 05/10/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use a aPI == 6UJ.L- [h(1 + :2:0)] -12k-ax ar r=ri (3) In equation (3), ~l and LO represent terms which result from velocity slip at the interface between the oil film and porous bearing. These terms can be shown to be expressed by [6) where, from Fig. 1, k 1/2/ a LO == -k-1/-2-/ a--'--+-h- L _ 3[2k + h k l/2 /a) 1 - h(k1/2/a + h) h == C[1 + E cos (\" + 0)). (4) (5) (6) Defining H == h/C and == k/C2, equations (4) and (5) can be written as where s LO ==- s+H 3(2<1> + sH) L1 == H(s + H) k 1/2 1/2 s==--==-- aC a (7) (8) (9) The dimensionless constant, a, depends on the porous material. For laminar channel flow, it has been estimated by Beavers, Sparrow, and Magnuson [7) that a is approximately 0.1. Since the fluid flow in po rous bearings can be assumed to be laminar and since the oil film thickness is small compared to the radius of the bearing, it is rea sonable to assume that an a of 0", + " (60) n=l Since the infinite series are truncated at n = Nand m = M, the series given by equation (60) will not be zero but it will have some finite value. Defining this value or truncation error for a given m as Om, equation (60) can be written as N L Cnm an(-1)n = Om n=I (61) The coefficients Cnm are obtained by choosing a l' which will minimize the sum of the squares of all the truncation errors, i.e., minimize G (1') where (62) 42 / VOL 101, JANUARY 1979 Load Capacity The load capacity of the bearing is found by integrating the pressure profile. Referring to Fig. 1, it can be seen that (L/2 (13 Wr = 2 Jo Jo P*I'; cos(7r - l' - 0) dO dz (63) (L/2 (13 Wt = 2 Jo Jo P*I'; sin (7r - l' - 0) dO dz (64) In dimensionless form, equations (63) and (64) can be, respectively, written as Wt = Wt C2 = (1 (flp*(O,Z)sin('Y+O)dOdZ 1';21lUL Jo Jo (65) (66) Substituting for p* (0, Z) given by equation (56), equations (65) and (66) can be written, respectively, as = = C Wr = L L --'!:!!!. (_I)m+1 n=l m=l ~m { [ COS (an - 1) {3 cos (an + 1) (3 a,,] X cos l' + - --- 2(an -1) 2(an + 1) (\u00a5,,2-1 . [Sin \u00ab(~n - 1) {3 sin (a\" + 1) {3]} + sm l' - 2(an - 1) 2(an + 1) (67) W t = f. f. _Cn_m (_1)m+1 {sin l' [_a_n_ n=l m=l ~m a n 2 - 1 _ cos (a\" - 1) {3 _ cos (an + 1) (3] 2(an - 1) 2(a\" + 1) [ sin (a\" - 1) {3 sin (an + 1) (3]) + cos l' - --'--\"--'-'-- 2(an - 1) 2(a\" + 1) (68) The total load is then Defining the Sommerfeld number as it is seen that (71) It can also be seen from Fig. 1 that the attitude angle, , is given by Coefficient of Friction The frictional force on the journal can be calculated from where aUI Th=llay y=h In dimensionless form, it can be shown that Th becomes _ IlU [H (1 1 \"\" ) ap*(O, Z) 1 - JJo] Th - - - + - \"1 + --- C23 ao H Using equation (56), equation (75) can be written as IlU [H ( 1 ) = = Th = - - 1 + - JJ1 L L Cnman C 2 3 n=l m=l 1 - JJo] X cos anO cos ~mZ + ~ (72) (73) (74) (75) (76) Transactions of the ASME Downloaded From: http://tribology.asmedigitalcollection" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002628_amm.373-375.38-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002628_amm.373-375.38-Figure1-1.png", + "caption": "Fig. 1. The experimental stand for research of the stiff rotor self-balancer on elastic supports: 1 \u2013 rotor; 2 \u2013 housing; 3 \u2013 rubber supports; 4 \u2013 engine; 5 \u2013 coupling; 6 \u2013 pendulums; 7 \u2013 additional load; 8 \u2013resistive strain gage; 9 \u2013 strain amplifier; 10 \u2013 oscillograph; 11 \u2013 tacho generator; 12 \u2013 stroboscopic tachometer; 13 \u2013 photo sensors", + "texts": [ + "199, Purdue University Libraries, West Lafayette, USA-02/06/15,21:43:46) treatments which could be called effect of \"crawling\" were found. In this case, the stiff rotor of the machine rotates with the operating speed, and the pendulums obtain the rotation frequency coinciding with one of the natural frequencies of oscillations of the rotor on elastic supports. The purpose of this research is studying the possibility of an effect occurring during the rotation of the stiff rotor on the elastic supports, having static and moment unbalance with four self-balancers of a pendular type. The scheme of experimental unit shown in Fig. 1 represents the massive rotor 1 mounted in the housing 2 on the elastic supports 3. Each of the supports represents a prismatic rubber element with the pressed plug where the rotor bearings are to be placed. The elastic elements are located on the housing with special metal holders. The stiffnesses of the supports in the horizontal and vertical directions are supposed to be equal. The rotor is driven to rotation by the direct current electric motor 4 by means of the coupling 5 allowing the misalignment of the rotation axes to be compensated", + " Each pendulum consists of the rolling bearing, the plug and the core with a carving for mounting loads to change the moments of pendulums inertia. The efficiency of balancing is estimated by the strain in the supports, and to measure them, there are strain gages 8 glued over the cross-pieces of the rotor supports being attached, the strain reinforcer 9 and the light-beam oscillograph 10 are used. Calibrating was carried out by means of the loading device and the dynamometer (they are not shown in Fig. 1). The rotation frequency of the pendulums and rotor was captured by photo sensors 13. The stroboscopic tachometer was used to supervise the location of pendulums and measurement of angular speeds of the rotor and pendulums 12. The mass and the moments of the rotor inertia, the stiffness of its supports, and the moments of the pendulums inertia are selected in such a way as to provide stability of the self-balancing motion mode. In particular, that is the condition according to which the angular speed of the rotor rotation has to be more than its critical speeds [9]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001298_s11431-010-3188-0-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001298_s11431-010-3188-0-Figure5-1.png", + "caption": "Figure 5 The positions of fore- and hind-feet on ceiling and wall (The angles between the first toe and other four toes are \u03b81, \u03b82, \u03b83 and \u03b84, in which \u03b85 is angle between the first toe and the direction of shear force).", + "texts": [ + " When geckos move on the ceiling or wall, the position of the foot is judged by the angle \u03b8 between line of toe T1 tip and toe T5 tip and gecko\u2019s locomotion direction. The angles \u03b8 of the fore- and hind-feet are (68.4\u00b17.4)\u00b0 (N=11) and (9.5\u00b18.5)\u00b0 (N=12) on ceiling, (69.9\u00b18)\u00b0 (N=11) and (8.1\u00b18.2)\u00b0 (N=12) on wall. The positions of fore-feet are no obvious difference on ceiling and wall as well as the hind-feet. The angle between FT s of toe T5 and resultant shear force of other toes is larger than 90\u00b0 ((118.1\u00b113.8)\u00b0, N=8), the angle between FT of toe T5 and resultant reaction force of other toes is larger than 90\u00b0, too ((142.5\u00b134.6)\u00b0, N=10) (Figure 5). Due to the adsorption capacity of the setae at the bottom of the toes, geckos can move freely on wall and ceiling. The physiological structure of gecko\u2019s toes is similar to zonal, which is suitable for being forced along toe radial. This is consistent with the measured results that the radial force of a toe is larger than tangential force of the toe. The angle \u03b2T is much smaller than the angle of seta desorption (about 30\u00b0 [12]), thus geckos can move with enough safe margins. When geckos move from wall to ceiling, more negative normal force must be generated to balance the gravity because the center of the gravity of external load changes relatively to the geometric position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002244_0022-4898(71)90025-5-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002244_0022-4898(71)90025-5-Figure2-1.png", + "caption": "FIG. 2. Stress state beneath a wheel - -appl ied stresses in the vertical plane: ~,,, r, principal stresses in the vertical plane due to applied stresses ~ , % ~ ~j.", + "texts": [ + " The third principal stress is perpendicular to the vertical plane of motion; this may be called lateral stress to distinguish it from the other stresses in the vertical plane. According to the Mohr-Coulomb theory, the failure plane is perpendicular to the plane in which the major and minor principal stresses lie. This is the vertical plane as long as the smaller principal stress in that plane is lower than the lateral stress; with an increase of the applied stresses, however, the lateral stress may become the minor principal stress, changing the ~1, ~3\" plane to an oblique one, as shown in Fig. 2. Just before this situation occurs and at the limit when the lateral stress and the smaller principal stress in the vertical plane are equal, failure conditions in the vertical plane govern the stresses; with lateral failure this condition still has to be maintained, otherwise the failure mode would revert to that in the vertical plane. Thus the effect of lateral failure is a limitation on the applied shear stresses; at this limit the stresses can be calculated on the basis of two dimensional failure conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000204_1.3002324-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000204_1.3002324-Figure3-1.png", + "caption": "Fig. 3 Photograph of the instrumented ball bearing", + "texts": [], + "surrounding_texts": [ + "w o d B i c d k\nw l c\nr s a a r t a r\nw t l e b\nw c m\nw a b\nF c\n0\nDownloaded Fr\nKN = B\nh 1\nhere h is the lubricant-film thickness, and B is the bulk modulus f the fluid. The bulk modulus can be expressed in terms of the ensity and longitudinal wave speed c of the fluid, such that = c2. Assuming that the media on either side of the layer have dentical acoustic properties both are steel in the rolling bearing ase considered in this paper and that the wave is normally incient the lubricant-film thickness can be extracted from the well nown quasistatic spring model as\nh = B fz R f 2 1 \u2212 R f 2 2\nhere z is the acoustic impedance of the media surrounding the ubricant-film, R f is the amplitude of the measured reflection oefficient, which is a function of the ultrasonic frequency f .\n2.2 Ball Bearing Lubrication. Figure 1 shows the configuation of a ball bearing outer-raceway and a ball in contact. As hown, the ball moves along the groove in the x-direction, the ball nd the groove are conformal in the y-direction, and the load is pplied in the z-direction. Subscripts a and b refer to the two olling elements i.e., the ball and the raceway, respectively . In his work, the regression equations of Dowson and Higginson 14 re used to predict the central film thickness hc in the ball-outeraceway contact.\nhc R = 2.69 U 0 E R 0.67 E 0.53 P E R 2 \u22120.067 1 \u2212 0.61e\u22120.73k\n3\nhere U is the mean surface speed, 0 is the lubricant viscosity at he contact entry, is the pressure-viscosity coefficient, P is the oad on the contact, k is the ellipticity ratio, E is the reduced lastic modulus, and R is the reduced radius of curvature given y\n1\nE =\n1 2 1 \u2212 a\n2\nEa +\n1 \u2212 b 2\nEb ,\n1\nR =\n1\nRax +\n1\nRbx +\n1\nRay +\n1\nRby 4\nhere E is Young\u2019s modulus, and is Poisson\u2019s ratio. The lubriated contact area is elliptical in shape with the minor a and ajor b semicontact radii given by\na = 6 PR kE 1/3 , b = 6k2 PR E 1/3\n5\nhere is the complete elliptical integral of the second kind. An pproximate relationship 15 for the maximum load on a single\nx\ny\nRbx Rby Rax Ray\nx\ny\n2a\n2b\nBall\nRaceway\nz\nig. 1 Configuration of ball bearing outer-raceway and ball in ontact\nall in a radially loaded ball bearing is used.\n11502-2 / Vol. 131, JANUARY 2009\nom: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/d\nP = 5W\nnb 6\nwhere W is the load on the whole bearing, and nb is the number of balls.\nThe inlet lubricant viscosity 0 was assumed to remain constant during these tests i.e., unaffected by the changes in bearing load and speed . The tests were performed relatively quickly data recorded within a 5 min interval so the bearing remained at room temperature throughout. Since there was no instrumentation to record the temperature of the oil and hence its viscosity at the contact entry, it was deemed preferable to use a constant value equal to that of the oil supply temperature.\n2.3 Lubricant Bulk Modulus. The bulk modulus used to determine the oil-film thickness from the layer stiffness Eq. 1 must be that of the oil under contact pressure. In EHL pressures, the oil compresses significantly and the local bulk modulus is several times higher than under ambient pressures. In this work, the compressibility model of Jacobson and Vinet 16 is used to determine the influence of pressure on the lubricant bulk modulus. They gave an equation of state to describe the behavior of the lubricant under pressure p.\np = 3B0\nx2 1 \u2212 x e 1\u2212x 7\nand the bulk modulus under pressure is given by\nB = B0\nx2 2 + \u2212 1 x \u2212 x2 e 1\u2212x 8\nwhere B0 is the bulk modulus at ambient pressure, is a lubricant specific parameter, and x is a function of the relative compression.\nx = 3 o\np 9\nwhere 0 is the density at ambient pressure, and p at pressure p. These equations are used for the oil both in its liquid state and after it undergoes solidification by employing different values of and B0. Values for these parameters were estimated for the test oil used in these experiments Shell Turbo T68 from generically similar oils tested by Jacobson and Vinet using a high pressure chamber. The bulk modulus for T68 at ambient pressure was determined from the speed of sound and density B= c2 of the oil and found to be B=1.84 GPa. Using Eqs. 7 \u2013 9 , the bulk modulus at a pressure of 1.5 GPa was found to increase to 21.2 GPa.\n2.4 Deformed Surface Geometry. Given the dimensions of the lubricated contact a, b, and hc, it is now possible to develop equations for the deformed geometries of the ball and outerraceway, both inside and outside the lubricated contact. The starting point is the geometry of the undeformed ball surface and outer-raceway. Taking the initial point of contact between the ball and the raceway as the origin, the geometry of the ball and raceway surfaces under no load is given by\nza = Rax 2 \u2212 x2 \u2212 y2 \u2212 Rax 10\nzb = Rbx 2 \u2212 x2 + Rby 2 \u2212 y2 \u2212 Rbx \u2212 Rby 11\nThe total elastic deflection at the central point of the contact i.e., at x ,y=0 is given by 17 .\nc = 4.5 nR P E 1/3\n12\nwhere is the complete elliptic integral of the first kind. The \u201cgap\u201d between the ball and raceway filled with lubricant in the contact region and either lubricant, air, or a mixture of the two outside this region is given by\nTransactions of the ASME\nata/journals/jotre9/28763/ on 03/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "w a e w r m l o\nw\nA b\nF b\nw\n3\no b i r t t c r a n i c T\nt\nJ\nDownloaded Fr\nh = hc + hg + e \u2212 c 13\nhere hg is the difference in the undeformed shapes of the ball nd raceway, i.e., hg x ,y =zb x ,y \u2212za x ,y and e is the total lastic deflection of the surfaces, i.e., e x ,y = a x ,y + b x ,y , here a and b are the elastic deflections of the ball and raceway, espectively. Because the ball and outer-raceway are typically ade from the same material, their elastic deflections are equal i.e., a= b . Assuming the oil-film thickness is the constant in the ubricated contact region, from Eq. 13 the elastic deflection of uter-raceway in the contact region is given by\na = b = e\n2 =\n1 2 c \u2212 hg for re 1 14\nhere\nre = x2 a2 + y2 b2\nHertzian contact model then allows the elastic deflection of the all and outer-raceway outside the contact region to be written as 18\na = b = c 2 \u2212 re\n2 sin\u22121 1\nre + re\n2 \u2212 1 for re 1 15\ninally, the deformed geometry of the ball and outer-raceway can e written as\nza,b = za,b \u2212 a,b 16 here a prime is now used to denote the deformed geometry.\nExperimental Apparatus and Procedure A piezoelectric thin film transducer was manufactured on the uter surface of the outer-raceway of a deep groove 6016 ball earing. Figures 2 and 3 show a schematic and photograph of the nstrumented bearing. Aluminum nitride AlN was deposited by adio frequency RF magnetron sputtering to form the piezoelecric film. The aluminum target was presputtered under a vacuum o remove the surface oxide. Nitrogen gas was then fed into the hamber and a film of AlN deposited on the bearing outeraceway. A coating thickness of approximately 4\u20136 m was chieved after several hours of sputtering. The substrate needed o preheating so there was no danger of the bearing steel temperng. Scanning electron microscope SEM examinations of the oating demonstrate that the film has a highly columnar structure. his gives them a strong piezoelectric property. An electrode was then deposited by evaporating aluminum\n2\nhrough a hole in a mask that measured 3 0.3 mm producing\nournal of Tribology\nom: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/d\nan electroded region of the same size. An electrical connection was then made between the electrode and a fine wire using conductive epoxy. The bearing raceway was grounded to form the other electrical connection. Finally, a protective shroud made from an electrically insulating material FR-4 PCB material was bonded to the outer-raceway to protect the electrode and the electrical connection during insertion of the bearing into the housing. More details of the coating process and transducer development are given in Ref. 19 .\nThe instrumented bearing was inserted into one of four bearing housings in a bearing test apparatus that supported an 80 mm diameter shaft see Ref. 20 for fuller details . This used a gravity feed system to lubricate the bearings with a mineral turbine oil Shell Turbo T68 . This apparatus enabled vertically upwards radial loads to be applied to the shaft and hence the ball at the top of the instrumented raceway was the most heavily loaded. Rotary shaft speed was controllable in the range 100\u20132900 rpm by a 7.5 kW electric motor. This control of load and speed then enabled control of the resultant oil-film thickness typically in the range 0\u20131 m .\nAn optical sensor was used, both to allow accurate triggering of the ultrasonic instrumentation and to measure shaft speed. This was triggered by reflective tape attached to the ball cage which rotates at half the shaft speed . When this tape passed the optical sensor it triggered a signal generator Agilent 33220A . After the addition of an adjustable delay the signal generator triggered a pulser-receiver Panametrics 5072PR at its maximum pulse repetition frequency, which was 20 kHz.\nThe pulses pass through the outer-raceway and reflect from the inner surface of the raceway below the transducer. If the ball is not located beneath the contact then the pulses are almost completely reflected back from the raceway due to the high reflectivity of a steel-air interface. When the trigger was set appropriately the pulses reflected back from the oil-film that forms between the ball and the raceway as it moves beneath the measurement location.\nThe pulser-receiver outputs a voltage spike containing frequencies centered on 200 MHz and had a receive bandwidth of up to 400 MHz. The reflected signals were then passed to a digital oscilloscope sampling frequency 5 GS/s and PC for storage and analysis. Figure 4 shows the reflected signal amplitude recorded for 500 successive data points the maximum capacity of the storage oscilloscope used . In this example, the data storage was triggered such that two ball passages can be seen.\nFigure 5 shows typical time and frequency domain plots of the signal reflected from the lubricant-film using the piezo thin film transducer on the bearing. In the time domain Fig. 5 a , a series of reverberations can be seen that, after experimentation, were found to be caused by electrical impedance mismatch between the\ncable and the receiver part of the pulser-receiver. To overcome\nJANUARY 2009, Vol. 131 / 011502-3\nata/journals/jotre9/28763/ on 03/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use", + "t t n d r u fi t 5 w o\nn r l r r i c\nF d h\n0\nDownloaded Fr\nhis problem a 3 m cable was used to enable the cable reverberaions to be separated and hence permit extraction of the first sigal, corresponding to the lubricant-layer reflection. Note that the istance from the transducer to the inner surface of the outeraceway distance was 4.5 mm and so the total return path of the ltrasonic echo was twice this. The time domain data of the oillm reflection shown in Fig. 5 a was converted, via a fast Fourier ransform, to the frequency domain. It can then be seen from Fig. b that the piezo thin film transducer is extremely wide band, ith energy in the range 50\u2013350 MHz and has a center frequency f around 200 MHz.\nThe reflection coefficient was measured by comparing the sigal reflected from the interface of interest to that from a known eference interface. In this case the interface of interest was the ubricant-oil-film between the outer-raceway and the ball and the eference interface was that from the inner surface of the outeraceway in the absence of both ball and lubricant i.e., a steel-air nterface and so Rref 1 . The reflection coefficient was then calulated from\n0 50 100 150 200 250 300 350 400 450 500 0\n0.2\n0.4\n0.6\n0.8\n1\n1.2\n1.4\n1.6\n1.8\n2\nR ef le ct ed si gn al am pl itu de (a rb .)\nData point number\nig. 4 Amplitude of the reflected signal for 500 successive ata points. The oil-film generation from two ball passages ave been captured.\n11502-4 / Vol. 131, JANUARY 2009\nom: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/d\nR f = Am f Aref f Rref 17\nwhere Am f is the amplitude of the signal reflected from the lubricant film layer, Aref f is the amplitude of the reference signal, and Rref is the reflection coefficient of the reference interface. The reflection coefficient calculated from Eq. 17 can then be used in Eq. 2 to extract the lubricant-film thickness assuming all other material constants, acoustic properties, and the reference reflection coefficient are known. The relevant acoustic properties used for bearing steel and the lubricant are given in Table 1.\n4 Results The reflection coefficient data obtained from the piezo thin film transducer under various load and speed conditions is shown in\nFig. 6. A distance of zero corresponds to the center of the lubricated contact. Away from the central region i.e., at distances greater than 300 m the reflection coefficient increases toward unity. Note that at distances of over 1 mm from the center of the lubricated contact the reflection coefficient remained within 10% of unity. Remote from the contact, signals are reflected from a steel-air or steel-oil interface depending on whether the bearing cavity between the balls remains partially filled or fully flooded with oil. The former would have a reflection coefficient R 1, the latter of R 0.95.\nWhen the contact is under the measurement location, the reflection coefficient reduces distinctly as a greater proportion of the signal is transmitted through the oil-film. It might be expected that since the oil-film in the contact region is largely parallel 17 that the reflection coefficient should remain at a constant level corresponding to a value predicted by Eq. 2 . However, it can be seen from Fig. 6 that the results particularly at W=2.5 and 5 kN appear to exhibit a local reflection coefficient maximum at the\n0 50 100 150 200 250 300 350 400 0\n0.5\n1\n1.5\n2\n2.5\n3\n3.5\n4\nFrequency (MHz)\nS pe ct ru m (a rb itr ar y un it)\nb)\nin film transducer deposited onto the bearing outer-raceway t lubricant-layer thickness and signals B are reverberations\nTable 1 Acoustic properties of lubricant and bearing steel\nDensity kg /m3\nLongitudinal wave velocity cl\nm/s Bulk modulus B GPa\nT68 at 0.1 MPa 876 1460 1.84 T68 at 1.5 GPa 1044 4500 21.2 Bearing steel 7900 5900 172\nth\nrac\nTransactions of the ASME\nata/journals/jotre9/28763/ on 03/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use" + ] + }, + { + "image_filename": "designv11_3_0001474_bf00533283-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001474_bf00533283-Figure1-1.png", + "caption": "Fig. 1. Geomet ry of the rotor", + "texts": [ + " Niordson for many valuable comments on the present paper. P. T. Pedersen: On Forward and Backward Precession of Rotors 27 which the natural frequencies do not appear in the boundary conditions. Also, in this case there is an infinite number of natural frequencies corresponding to forward synchronous precession. However, it should be emphasized tha t the effect of the transverse shear force does not influence the results obtained concerning the excitation of forward and backward precession. 2. Basic Assumptions The rotor model is shown in Fig. 1 in a fixed system of rectangular coordinates. The rotor consists of a nonuniform, continuous circular shaft of lenght L supported at the ends by two anisotropic flexible bearings (I and II). We will assume tha t the shaft is subjected to rotat ing internal viscous damping forces and s ta t ionary external damping forces which oppose rotat ion and translation of the cross-sections of the shaft. The properties of the flexible bearings are such tha t the displacement components x, y from the neutral position produce forces Px and Py at the journal where for bearing I : ~DI x ~ S~I X(O) \"~ S~2 y(O) , -Ply ~ SI1 X(O) ~- SI22 y(o) where s~2 = s~l" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002505_12.918523-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002505_12.918523-Figure3-1.png", + "caption": "Fig. 3 Interception for minimal impact to the base satellite based on the contact force direction", + "texts": [ + " This is very difficult or impossible when the target satellite has a fast tumbling motion because the tip speed of a manipulator is always limited not only by the joint rate limits but also by the attitude tolerance of the servicing satellite. In such a case, a strategy using the second condition as its control goal becomes more attractive because it does not requires zero relative velocity, but such an approach has not been studied in the past Therefore, we will focus our study on achieving the second condition. Proc. of SPIE Vol. 8385 83850J-5 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 11/02/2015 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx As shown in Fig. 3, the second condition means that the angle \u03b2 (between the relative velocity and the position vector of the grasping handle of the target satellite) should be zero. In such a case, the major component of the impact force (assuming mainly along the relative-velocity direction) will pass through the mass center of the servicing system and thus, cause no angular moment to the servicing satellite. Of course, this is only an ideal case. In a general tumbling case, the direction of the relative velocity may never pass through the mass center of the servicing satellite" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001823_j.triboint.2009.10.015-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001823_j.triboint.2009.10.015-Figure3-1.png", + "caption": "Fig. 3. Photo of HAFB with bump foils: (a) Overall concept, (b) orifice and thermocouple locations and (c) photo of HAFB with bump foils.", + "texts": [ + " Since the rotor is supported on ball bearings and there is no significant external damping, undamped critical speeds were estimated. The first critical speed for the forward whirl was estimated as 43,325 rpm with predicted ball bearing stiffness of 1.75 108 N/m. First bending mode was predicted as over 202,500 rpm, far above the maximum operating speed of the test rig. The mode shape of the shaft at the first critical speed is shown in Fig. 2. A schematic description of HAFB with two bump foil strips and single continuous top foil is shown in Fig. 3(a) However, manufactured HAFB shown in Fig. 3(c), has four bump foil strips with identical stiffness along axial direction (essentially Generation I). Even though the HAFB could be constructed to Generation II or III bearings, current focus is on Generation I to see the effectiveness of the hybrid operation. Externally pressurized air was supplied through four steel feed tubes (OD: 1.27 mm/ID: 0.97 mm), which are connected to the backside of the top-foil using silicone rubber tubing (OD: 1.65 mm/ID: 0.76 mm) as shown in Fig. 3(c). The rubber tubing provides flexibility and is easier to glue on to the curved top foil surface. Circumferential locations of the orifices are at y=721, 1661, 2471, and 3411 as shown in Fig. 3(b). The purpose of the unsymmetrical placement of the orifices is to allow bump foils not to lose sliding contacts with the bearing sleeve (see detailed view of orifice location in Fig. 3(a). One thermocouple was attached to the back side of the top foil at downstream of the second feed tube as shown in Fig. 3(b). The top foil has Teflons coating but rotor was a bare stainless steel. Parameters of the HAFB are presented in Table 1. The bearing nominal clearance in the table is an estimated value from precise measurements of bearing sleeve inner diameter, top foil thickness, bump foil thickness, bump foil height, and rotor outer diameter. The bump stiffness was calculated using the formula for free\u2013free case presented by Iordanoff [17]. The experimental setup is shown in Fig. 4(a). The load is applied to the bearing using a pulley mechanism as depicted in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003907_iecon.2013.6699627-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003907_iecon.2013.6699627-Figure8-1.png", + "caption": "Fig. 8. Misaligned dq-frame [8].", + "texts": [ + " Based on the superposition law, the high frequency part is separated as[ 0 0 ] = [ Ldd Ldq Lqd Lqq ] p [ iedh ieqh ] + \u03c9h [ Ldf Lqf ] Ih sin\u03c9ht (9) In deriving (9), the terms, ridh, riqh, \u2212\u03c9e\u03bb e q , and \u03c9e\u03bb e d are neglected since \u03c9h is much larger than \u03c9e and \u03c9hLdd \u226b rs, and \u03c9hLqq \u226b rs. Therefore, the solution to (9) is obtained such that[ iedh(t) ieqh(t) ] = Ih [ \u03b1A \u03b1B ] cos\u03c9ht+ [ iedh(0) ieqh(0) ] . (10) where \u03b1A = LqqLdf \u2212 LdqLqf LddLqq \u2212 LdqLqd , \u03b1B = \u2212LqdLdf + LddLqf LddLqq \u2212 LdqLqd . Note that (10) is an equation in the synchronous(rotor) frame based on the right angle, \u03b8e. The angle error is defined as \u2206\u03b8e = \u03b8\u0302e \u2212 \u03b8e, (11) where \u03b8\u0302e is an estimation of the real position \u03b8e in Fig. 8. To obtain the current in an estimated frame, we transform (10) such that[ i\u0302edh i\u0302eqh ] = [ cos\u2206\u03b8e sin\u2206\u03b8e \u2212 sin\u2206\u03b8e cos\u2206\u03b8e ] [ iedh ieqh ] = Ih [ \u03b1A cos\u2206\u03b8e + \u03b1B sin\u2206\u03b8e \u2212\u03b1A sin\u2206\u03b8e + \u03b1B cos\u2206\u03b8e ] cos\u03c9ht = Ih \u221a \u03b12 A + \u03b12 B [ cos(\u03b4\u03b8\u2032e \u2212\u2206\u03b8e) \u2212 sin(\u03b4\u03b8\u2032e \u2212\u2206\u03b8e) ] cos\u03c9ht. (12) where \u03b4\u03b8\u2032e = tan\u22121( \u03b1B \u03b1A ) = tan\u22121 ( \u2212LqdLdf + LddLqf LqqLdf \u2212 LdqLqf ) . Note that (12) contains functions of \u2206\u03b8. In contrast, functions of 2\u2206\u03b8 are derived if the signal is injected to the stator windings. Hereforth, the common synchronous rectification method is utilized to extract \u2206\u03b8 information" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001578_jnn.2009.se41-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001578_jnn.2009.se41-Figure7-1.png", + "caption": "Fig. 7. Schematic of the LCPCF SERS sensor and its cross-sectional view. The spectrometer above the surface contains a CCD detector, a monochromator, and electronics for data collection. Reprinted with permission from [23], Y. Zhang et al., Appl. Phys. Lett. 90, 193504", + "texts": [ + "66\u201368 By changing the fusion power, fusion duration and position of 2242 J. Nanosci. Nanotechnol. 9, 2234\u20132246, 2009 Delivered by Ingenta to: Purdue University Libraries IP: 194.50.116.218 On: Thu, 09 Jun 2016 07:37:20 Copyright: American Scientific Publishers Shi et al. Molecular Fiber Sensors Based on Surface Enhanced Raman Scattering (SERS) Institute of Physics. the fiber, the two tips of the fusion splicer can generate heat uniformly and seal the cladding holes more exactly than the ordinary flame. The schematic of the LCPCF sensor is presented in Figure 7, in which the laser excitation light is coupled into one end of the HCPCF tip. The other side of the HCPCF tip is dipped into the SERS/analyte solution, and the SERS signal is transferred back to the Raman spectrometer through the same fiber. For biosensing application, a diode laser with 785 nm excitation wavelength is (2007). \u00a9 2007, American Institute of Physics. selected to avoid burning of biomolecules. With the tip of the PCF dipped into the solution, the capillary effect will force the solution into all the hollow regions of the PCF" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001294_acc.2010.5531005-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001294_acc.2010.5531005-Figure1-1.png", + "caption": "Fig. 1. Quadrotor helicopter configuration frame system", + "texts": [ + " In Section IV, extended experimental results that prove the efficacy of the proposed scheme are presented followed by the conclusions in the last section V. II. QUADROTOR HELICOPTER MODELING The model of the UqH utilized in this work, assumes that the structure is rigid and symmetrical, the center of gravity and the body fixed frame origin coincide, the propellers are rigid and the thrust and drag forces are proportional to the square of propeller\u2019s speed. The electro\u2013mechanical structure of the UqH under study, and the relative coordinate systems are presented in Figure 1. The UqH\u2019s nonlinear dynamics [14] is characterized by a set of twelve high non\u2013linear state equations in the form: X\u0307 = f (X,U)+W (1) with f a non\u2013linear function, W corresponds to the additive effects of the environmental (wind) disturbances, X the state vector, and U the input vector, where: X = [\u03c6 \u03c6\u0307 \u03b8 \u03b8\u0307 \u03c8 \u03c8\u0307 z z\u0307 x x\u0307 y y\u0307] (2) U = [U1 U2 U3 U4 \u03a9r] . (3) 978-1-4244-7427-1/10/$26.00 \u00a92010 AACC 4451 The control inputs in (1) are produced by the following combinations of the angular speeds of the four UqH\u2019s rotors as: U1 = b(\u03a92 1 +\u03a92 2 +\u03a92 3 +\u03a92 4) U2 = b(\u2212\u03a92 2 +\u03a92 4) U3 = b(\u03a92 1 \u2212\u03a92 3) U4 = d(\u2212\u03a92 1 +\u03a92 3 \u2212\u03a92 3 +\u03a92 4) \u03a9r = \u2212\u03a91 +\u03a92 \u2212\u03a93 +\u03a94 where b is the thrust coefficient, and d is the drag coefficient, while the input U1 is related with the total thrust and the inputs U2, U3, U4 are related with the rotations of the quadrotor and \u03a9r is the overall residual angular velocity of the motors" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003113_robio.2011.6181524-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003113_robio.2011.6181524-Figure1-1.png", + "caption": "Fig. 1. Prototype of RT-Walker (a), and the rear wheel and servo brake system (b)", + "texts": [ + " The experiments done to confirm the validity of the method are included in VII. The conclusions would be the last part of this article. II. RT WALKER Inspired by the proposed concept of passive robotics by Goswami et al. [11], we have used servo brakes to control the motion of a passive-type walker. Since there is no chance for unintentional movements of the system, passive robots are intrinsically safe and suitable for the systems having physical interaction with humans. The developed RT Walker is a prototype with a support frame, two passive casters, and controller (figure 1). Unlike other passive robots, RT Walker doesn\u2019t have any servo motors to control the servo brakes. The rear wheels are equipped with powder brakes which can be used to change the brake torques of rear wheels of RT Walker according to input current. The goal is to activate the brakes of RT-Walker in the situations where the user is about to fall or the user is trying to stand up or sit down. The control framework of the system is shown in figure 2. Here the main focus is on detecting any non-walking state of motion for which the response of the system would be merely stopping" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003105_j.ijsolstr.2011.09.017-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003105_j.ijsolstr.2011.09.017-Figure3-1.png", + "caption": "Fig. 3. Isochoric constrained Ziegler column problem.", + "texts": [ + " When the load p is higher than ps 2, both K and Ks admit two negative eigenvalues. Whatever the vector x is, the quantity tx K x is negative. The unique equilibrium solution h 1 \u00bc 0; h 2 \u00bc 0 is locally unstable, and any velocity disturbance applied to h1 and h2 provokes an increase in the kinetic energy of the system. The next subsection investigates the fundamental role played by external constraints in the local destabilization of the system. The generalized Ziegler column problem is considered by henceforth preventing the lateral deviation of point C (Fig. 3). This constraint reads: h1 \u00fe h2 \u00bc 0 \u00f050\u00de Interestingly, Eq. (50) corresponds to a so-called isochoric condition, like that used in soil mechanics for the standard undrained triaxial test: an axial loading is applied to the soil specimen, while the isochoric condition (the volume is maintained constant) _e1 \u00fe 2 _e3 \u00bc 0 is prescribed ( _e2 \u00bc _e3 for axisymmetric reasons). Assume that the system is initially loaded so that p > ps 1. The equilibrium configuration \u00f0h 1 \u00bc 0; h 2 \u00bc 0\u00de is locally unstable" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002295_0954410011433237-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002295_0954410011433237-Figure1-1.png", + "caption": "Figure 1. Schematic of pressure-balanced pintle-type gas regulating system.23", + "texts": [ + " Mathematical model of the ducted rocket The system is described according to the dominant physical phenomena. The typical controlled plant for thrust control includes an actuator (i.e. the gas regulating system) and the model of ducted rocket, which will be separately introduced in the following paragraphs. In addition, for the inlet buzz protecting control, it should also include the part of inlet. Gas regulating system. A pressure-balanced gas regulating system, as given in the study of Shi et al.,23 is depicted in Figure 1. In the diagram, when the gas flow demand rises, the servo valve (2) can make the gas go into the valve head chamber (6) from a gas bottle (1). Then, the champer pressure is increased and the valve head (5)\u2019s force balance is broken. Meanwhile, the valve head (5) goes forward along the valve stem (7), which makes the gas generator\u2019s (3) pressure rise but its throat area (4) decrease. The reader can refer to the study of Shi et al.23 for further details. The mathematical model of gas flow regulating system for ducted rocket can be described by the following three equations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003538_aim.2011.6026975-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003538_aim.2011.6026975-Figure5-1.png", + "caption": "Fig. 5. Minimum jerk model", + "texts": [ + " When a robot approaches a human, it is necessary for the robot\u2019s hand to have an acceptable trajectory in which the velocity and acceleration change smoothly. Shibata [5] discussed that the bell-shaped velocity pattern is acceptable for human emotions when a robot approaches a human. On the other hand, the minimum jerk model [6], which minimizes equation (1), can accurately reproduce the pointto-point motion of human hands. Here, (x, y) is the position of a human hand, and Tf is the movement time. In this model, the acceleration has smooth changes, and the velocity pattern has a symmetrical bell-shaped profile, as shown in Fig.5. Thus, it is expected that the trajectory of the robot, which is preferred by humans, is generated using the minimum jerk model. Therefore, in this paper, the handshake request motion of the robot is generated according to the minimum jerk model. C = 1 2 \u222b Tf 0 ( ( d3x dt3 )2 + ( d3y dt3 )2 ) dt (1) In the minimum jerk model, the robot velocity V (t) is calculated from the maximum velocity Vmax and the movement time Tf by equation (2). The maximum velocity Vmax is calculated by equation (3). Here, r0 and rf are the initial and target angles, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003228_s11042-013-1359-2-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003228_s11042-013-1359-2-Figure7-1.png", + "caption": "Fig. 7 Wrist angle computation from the two rotation matrices of the IMU sensors on the left forearm and the golf club", + "texts": [ + " However, this ZYX Euler angles-based method has the singularity problem. This means that in some cases, there exist infinitely many choices of \u03b1, \u03b2, and \u03b3 for the given R(\u03b1, \u03b2, \u03b3). It has been proven mathematically that there exists no single parameterization of the rotation that is free from singularity. Therefore, instead of the ZYX Euler angles-based method, the proposed system calculates the wrist angle by using only two x vectors of Rarm(t) and Rclub(t), which are aligned along the direction of the forearm and golf club. Figure 7 illustrates the wrist angle computation. Let xarm(t) and xclub(t) be the x vectors of Rarm(t) and Rclub(t) at time t, respectively. We can calculate the wrist angle \u03b8(t) as \u03b8\u00f0t\u00de \u00bc acos xarm\u00f0t\u00de xclub\u00f0t\u00de\u00f0 \u00de; 0 \u03b8\u00f0t\u00de p: \u00f05\u00de The system outputs the wrist angle and stores not only the wrist angle but also the synchronized set of x vectors from the forearm and the golf club. Data are collected sequentially and stored into memory as an array data type until the end of the user\u2019s swing, at which point they are used to analyze the uncocking motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000540_insi.2008.50.4.195-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000540_insi.2008.50.4.195-Figure3-1.png", + "caption": "Figure 3. Vectors of gear condition with modified, unmodified and pitted gears: W, 1, 2 (class 6 PS) \u2013 modified gears; 3, 4, 5 \u2013 unmodified gears (class 8, PS), 6, 7, 8 \u2013 unmodified gears (class 7, PS); 9 \u2013 unmodified gears (class 6, PS); PS \u2013 Polish", + "texts": [ + " There are different starting assumptions taken for the condition monitoring. They can be described as is given in this paper in the chapter about factor analysis. As will be shown, usually, only some factors presented in Figure 1 and Figure 2 are taken into consideration. Referring to the literature[4][21] it can be seen that the aim of the presented diagnostic method is the diagnostic assessment of the gearbox condition described by the limited imperfections given by the dimension and shape deviations (according to Polish Standards (PS) similar to ISO). Figure 3 presents the vectors of the gear condition with the modified, unmodified and pitted gears: W, 1, 2 (class 6 PS) \u2013 modified gears; 3, 4, 5 \u2013 unmodified gears (class 8, PS); 6, 7, 8 \u2013 unmodified gears (class 7, PS); 9 \u2013 unmodified gears (class 6, PS); 10-pitted gear. For the condition evaluation, the coherence gearbox condition monitoring method (CGCMM) is used as described in the literature[4]. The gear condition assessment is obtained by measurement of six coherence components for a gearing frequency (900 Hz) and its harmonics (1800-5400 Hz). The signals are received from the four points of a gearbox housing. From the four points one can obtain six independent measurements of the coherence function. The values of squared coherence function components are averaged and treated as the vector components of six-dimension space. The magnitudes of the condition vectors are given in Figure 3. CGCMM were developed after the statement that the vibration meshing spectrum components of the signal are the measure of a gear condition. Using Figure 1 as the background for revealing vibration signal properties, it can be concluded that the vibration spectrum components can be the measure of the gearbox condition. On the basis of such design factor as the number of gear teeth and the operation factor as the shaft rotation frequency, one can calculate the meshing frequency and by its multiplication by constants (1-6) one can calculate the meshing frequency components", + " It is presumed that the energy of the white noise is constant but with increasing the gear dimension deviations and occurring the distributed faults, the energy of the meshing components increases, and the measure of the coherence components are the measure of the gearing condition. Taking the first six components as vector Standard similar to ISO, 10- pitted gear 198 Insight Vol 50 No 4 April 2008 components in six-dimension space, the length of the vector and the relative angles between the reference vector are calculated. The vector magnitudes for the gearing in different conditions are given in Figure 3. One of the problems is the number of components that should be taken into consideration by the gearing condition evaluation. This problem is discussed in the literature[21] when the principal component analysis (PCA) is taken. The result of PCA is given in Figure 4. Figure 4 shows that all six features components should be used. Using six components, as shown in Figure 4, 100% of information is conveyed. The above consideration proves that only some factors described by Figure 1 are taken into consideration when the diagnostic method is developed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000233_tie.2007.898297-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000233_tie.2007.898297-Figure4-1.png", + "caption": "Fig. 4. Fabrication process. (a) Original SOI wafer. (b) DRIE to etch airnozzle at front-side wafer. (c) Flipped wafer and DRIE etching of actuation microstructures at back-side wafer. (d) Flipped wafer and HF vapor to release actuation microstructures.", + "texts": [ + " Illustrations give views of distributed airflow surface with object, and focus on front-side and back-side of the single pneumatic microactuator. At the bottom of the substrate, a mobile microvalve is supported by suspension beams and actuated by electrostatic effect. Two electrodes generate electrostatic force are aligned next to the mobile microstructure part and fixed to the top substrate. The microvalve moves left or right by applying the electrostatic force. Airflow from a simple nozzle passes from the back-side of the microactuator. Fig. 4 shows top and cross section-views of two main steps of the batch fabrication process based on an SOI wafer (100/2/100 \u00b5m), as shown in Fig. 4(a). In the first step of the actuator fabrication process, the air-nozzle is made out of the silicon base substrate by the deep reactive ion etching (DRIE) process with photoresist mask, as shown in Fig. 4(b). Second, the SOI wafer is flipped and the back-side layer is etched into the microactuators by the DRIE process with photoresist mask, as shown in Fig. 4(c). Then, the SOI wafer is flipped again, and the microactuators are released by sacrificial layer (BOX layer) etching using HF vapor (hydrofluoric acid) as presented in Fig. 4(d). HF vapor play a significant role in avoiding sticking problem and realize freestanding structures. This process has been reported in previous work [22]. Fig. 5 presents fabrication results by scanning electron microscope (SEM) images of both sides of the device. Fig. 5(a) shows an image of the final MEMS chip and its packaging for control and voltage driving. The size of the device is about 35 \u00d7 35 mm2 for 560 MEMS-based pneumatic microactuator array. Fig. 5(b) is the SEM image the front-side of the airflow surface device, an array of horizontal and vertical nozzles" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000832_j.jfranklin.2008.05.006-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000832_j.jfranklin.2008.05.006-Figure1-1.png", + "caption": "Fig. 1. The dynamics of planar V/STOL (PVTOL) aircraft.", + "texts": [ + " In order to allow lateral maneuverability in the jet-borne operation, the aircraft also has a reaction control system (RCS) to provide moment around the aircraft center of mass. In the case that the bleed air from the reaction control valves produces force which is not perpendicular to the pitch axis, there will be a coupling effect between the angle rolling moment and lateral moving force. By restricting the aircraft to the jet-borne operation, i.e., thrust directed to the bottom of the aircraft, we have simplified dynamics which describes the motion of the aircraft in the vertical\u2013lateral directions, i.e., a planar V/STOL (PVTOL) aircraft as shown in Fig. 1. The aircraft states are the position of center of mass, (X, Y), the roll angle y, and the corresponding velocities, \u00f0 _X ; _Y ; _y\u00de. The control input is the thrust directed to the bottom of aircraft U1 and the moment around the aircraft center of mass U2. Let the quantity of lateral force induced by rolling moment be denoted by e0, then we have the aircraft ARTICLE IN PRESS S.-L. Wu et al. / Journal of the Franklin Institute 345 (2008) 906\u2013925 909 dynamics written as m \u20acX \u00bc sin yU1 \u00fe 0 cos yU2; m \u20acY \u00bc cos yU1 \u00fe 0 sin yU2 mg; J \u20acy \u00bc U2; 8>< >: (1) where mg is the gravity force imposed in the aircraft center of mass and J is the mass moment of inertia around the axis through the aircraft center of mass and along the fuselage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002186_s11740-010-0289-3-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002186_s11740-010-0289-3-Figure7-1.png", + "caption": "Fig. 7 Hydraulic design of the GIC", + "texts": [ + " In contrast to that, the input parameters for the hydraulic design of the GIC (i.e. the curvature of the channels) are the wheel diameter, the cross-sectional area of the coolant outlets (as presented in chapter 3.2), and the desired grinding wheel speed (here: up to 100 m/s). The criterion used for the curvature of the channels is the so called \u2018\u2018shock-free entry\u2019\u2019 [14]. This means that the angle of the flow velocity at the beginning of the channels (b1) has to be the same as the angle of the turbine blade at the beginning of the channels (see Fig. 7). In this case, the turbine blades are the walls of the channels. The angle of the flow velocity depends on the flow rate, the cross-sectional area of the channels and the rotational speed (in this case: the cutting speed). The angle of the turbine blade at the end of the channels (i.e. at the circumference of the wheel) determines the possible delivery head. While common centrifugal pump impellers provide angles of b2 = 20\u201325 , for the GIC an angle of b2 = 45 was set. First of all, this results in easier to manufacture undercuts in the transition area of the channels", + " Furthermore it results in less fluid friction inside the channels and higher fluid pressures at the outlets. With the angles b1 and b2, the curvature of the channels can be determined. This can be done analytically or with the use of computer-aided design (CAD). For the determination via CAD, the position of the beginning and the end of a channel within the GIC is set. Then the angle b1 at the beginning of the channel and the angle b2 at the end of the channel are set. With these geometrical conditions, only one possible radius for the curvature of the channel exists (see Fig. 7) [16]. The same procedure was done for the impeller that covers the mounting-screws to assure a shock-free entry of the cooling lubricant to the channels (see chapter 3.2). After the analytical specification of the hydraulic design is determined, computational fluid dynamics (CFD) simulations were used to gain detailed information about the fluid flow through the cooling channels and hence the hydraulic efficiency. Computational fluid dynamics, which is based on Navier\u2013Stokes-equations, simulates pressure and flow velocities for discrete points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002362_tmag.2012.2199094-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002362_tmag.2012.2199094-Figure1-1.png", + "caption": "Fig. 1. Illustration of a permanent magnet mover of a rotary-linear actuator with a checkerboard magnetization with representing the axial force, and representing the torque.", + "texts": [ + " Various configurations of rotary-linear actuators are described in literature based on classical induction-, switched reluctance-, and permanent magnet machines [1]\u2013[3]. This paper focuses on the class of PM-actuators. In [4], [5], actuators are presented with a single magnetization pattern on the mover and a set of coils to provide the two directions of motion. The magnetization in these actuators has a checkerboard orientation, i.e., magnets with opposing polarity are alternatingly placed in the axial- and circumferential direction, as illustrated in Fig. 1. The variation of the magnetic field of the PMs in the circumferential, -, direction is used to produce torque, , while the magnetic field variation in the axial, -, direction is used to produce force, . The actuator in [4] has a mover with a checkerboard magnetization pattern as shown in Fig. 2(a). The magnetization pattern covers 50% of the mover surface with magnets. Furthermore, at a fixed , the magnet polarization is constant at all circumferential positions and similarly, at a fixed , the magnet polarization is constant at all axial positions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure7-1.png", + "caption": "Fig. 7. (a) A double chain linkage consisting of four non-intersecting pairs. (b) The edge vectors and the inclination angles.", + "texts": [ + " (6), where a \u00bc p1j j cos 0\u00fe p2j j cos a1 \u00fe p3j j cos\u00f0a1 \u00fe a2\u00de \u00fe p4j j cos\u00f0a1 \u00fe a2 \u00fe a3\u00de; \u00f018\u00de b \u00bc p1j j sin 0\u00fe p2j j sin a1 \u00fe p3j j sin\u00f0a1 \u00fe a2\u00de \u00fe p4j j sin\u00f0a1 \u00fe a2 \u00fe a3\u00de \u00f019\u00de and c \u00bc q1j j cos 0\u00fe q2j j cos b1 \u00fe q3j j cos\u00f0b1 \u00fe b2\u00de \u00fe q4j j cos\u00f0b1 \u00fe b2 \u00fe b3\u00de: \u00f020\u00de The same applies to the other projection equation. Moreover, all of these coefficients must be zero in order for the equations to hold for any h, which, together with Eqs. (13) and (14), have been found to be equivalent to vectors p1, p2, p3, p4 and q1, q2, q3, q4 forming closed loops, respectively, i.e., p1 \u00fe p2 \u00fe p3 \u00fe p4 \u00bc 0 \u00f021\u00de . 6. (a) Pieces forming a non-intersecting pair; and (b) a double chain linkage with four non-intersecting pairs prior to forming a closed loop. and q1 \u00fe q2 \u00fe q3 \u00fe q4 \u00bc 0; \u00f022\u00de see Fig. 7. Alteration of rotation angle h only causes the rotation of the closed vector loops as a whole. This proof can be easily extended to double chain with any even number of non-intersecting pairs. Although the expressions of Eqs. (21) and (22) looked similar to what we obtained for the double chain with intersecting pairs, they are in fact different. This is due to the way the pieces were arranged under the parallelogram condition. Here p1, p2, p3 and p4 represent the lengths of opposite pieces, see Fig. 7. So do q1, q2, q3 and q4. Now insert an additional pair into the assembly to make a double chain consisting of five non-intersecting pairs, see Fig. 8a. Note that, unlike the previous double chain with four non-intersecting pairs, the pair on the left consists of a piece with edge lengths of p5 and q1 whereas the other piece has edge lengths of q5 and p1 in order to preserve the loop parallelogram condition. The angles sustained by p5 and q1, and by q5 and p1, are represented by a5 and b5, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.34-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.34-1.png", + "caption": "Fig. 2.34 The model of the Daniels Co-Axial Escapement", + "texts": [ + " In the past century, the design of the escapement has continued to evolve. The most significant invention is perhaps the Daniels co-axial double-wheel escapement. It is the masterpiece of Dr. George Daniels (1926\u2013) (Wikipedia 2009). Dr. Daniels is a professional horologist with many achievements. Besides inventing the co-axial escapement, he is also the author of several books on mechanical watch movement (Cecil Cluttoh and George Daniels 1979; George Daniels 1981; George Daniels 2011) and the former president of the Horological Institute. Figure 2.34 shows the model of the Daniels co-axial escapement. It is more complicated than the Swiss lever escapement and has three levels. Figure 2.35 shows the three levels of the escapement. On Level 1, the balance wheel contacts the pallet fork. The guard pins are also on this level. The escape wheel has two levels, one for the inner escape wheel and the other for the outer escape wheel, with 12 teeth on each level. On Levels 2 and 3, the two levels of the pallet fork contact the two levels of the escape wheel" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003288_1.4769758-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003288_1.4769758-Figure5-1.png", + "caption": "FIG. 5. (a) Cylinder floating on a curved interface in the presence of gravity. (b) Gravity resolved into normal and tangential components for a each cylinder.", + "texts": [ + " Hu and Bush49 have shown that tiny beetle larvae use curvature gradients to propel themselves on a fluid meniscus. Cavallaro et al.50 use curvature gradients to migrate rod-like particles with the aim of creating complex particle rafts. The contact angle in their experiments was 120\u25e6 and the particles aligned perpendicular to the direction of motion. In this section, we develop an asymptotic theory valid for small particles floating on arbitrary curved interface and our results have direct relevance to the experiments discussed above. Consider a curved interface as shown in Figure 5(a), where \u03b7(\u03be ) is the height of the interface from a known reference. The background curvature provides a capillary pressure contribution to the pressure balance on the interface. The presence of particles distorts the background interface modifying the local curvature around the particles. The gravitational force can be resolved along and perpendicular to the background interface as shown in Figure 5(b), where \u03b3 is the angle made by the normal to the undisturbed interface with the vertical. We assume that the radius of the cylinders is small compared to the characteristic length scale of the background, the latter being typically O(lc). This is equivalent to the Bond number being small, and hence the interface distortion due to the particle will be small compared to capillary length. In this limit, we can linearize the interface deformation about the undistorted curved interface and derive a simple expression for the perturbed curvature of the interface as shown in the Appendix", + " The perturbed curvature is in turn used to derive This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to IP: 142.157.129.89 On: Thu, 27 Nov 2014 17:06:10 a perturbed Young-Laplace equation which can be written in dimensional form as hxx + a(\u03be )hx + b(\u03be )h = 0, (40) where (x, y) is the local coordinate system defined along and perpendicular to the interface lying along the undisturbed interface as shown in Figure 5(a), a and b are parameters characterizing the background interface. Due to the small size of the particles, we calculate a(\u03be ) and b(\u03be ) at the origin of the local coordinate system, x = 0. This allows us to obtain a simple solution to the above equation instead of a power series solution. The general solution takes the simple form h(x) = c1e\u03bb\u2212x + c2e\u03bb+x , (41) where the relation between the eigenvalues, \u03bb\u00b1, and interface shape parameters like curvature and slope are given in the Appendix. Due to the curvature of the background interface, the interface distortion is not symmetric on either side of the particle, leading to a lateral force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001095_s11044-010-9236-5-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001095_s11044-010-9236-5-Figure4-1.png", + "caption": "Fig. 4 Joint coordinate systems", + "texts": [ + " For this reason, the orthonormal triad can be defined using these two vectors as: Aik s ( eik ) = [ iiks jiks kik s ] (13) where the direction cosine vectors are defined using Gram\u2013Schmidt orthogonalization process as iiks = jiks \u00d7 kik s , jiks = \u2202rik/\u2202yi |\u2202rik/\u2202yi | , kik s = \u2202rik/\u2202zi \u2212 hik\u2202rik/\u2202yi |\u2202rik/\u2202zi \u2212 hik\u2202rik/\u2202yi | (14) and hik = \u2202rik/\u2202zi \u00b7 \u2202rik/\u2202yi \u2202rik/\u2202yi \u00b7 \u2202rik/\u2202yi . (15) It is important to note that vector iiks is not necessarily tangential to the beam centerline due to deformation of cross section as shown in Fig. 4(a). Such joint coordinate system is called cross section frame [5]. It is also important to note that the orientation matrix obtained using (13) is different from the orientation matrix obtained using the Polar Decomposition Theorem given by (12). If deformation of cross section remains small, one can use a tangent frame which takes a simpler form of the orientation matrix as compared to the cross section frame [5]. In this case, one has Aik t ( eik ) = [ iikt jikt kik t ] (16) where the direction cosine vectors are defined using the gradient vectors \u2202rik/\u2202xi and \u2202rik/\u2202yi as iikt = \u2202 r\u0302ik/\u2202xi, jikt = kik t \u00d7 iikt , kik t = \u2202 r\u0302ik/\u2202xi \u00d7 \u2202 r\u0302ik/\u2202yi |\u2202 r\u0302ik/\u2202xi \u00d7 \u2202 r\u0302ik/\u2202yi | (17) where \u2202 r\u0302ik/\u2202xi and \u2202 r\u0302ik/\u2202yi are unit vectors of \u2202rik/\u2202xi and\u2202rik/\u2202xi , respectively. As shown in Fig. 4(b), X-axis of the tangent frame, which is defined by a unit vector iikt in (17), is always tangential to the beam centerline, and this vector is given as iikt = \u2202rik \u2202sik = \u2202rik \u2202xi ( \u2202sik \u2202xi )\u22121 = \u2202rik/\u2202xi |\u2202rik/\u2202xi | (18) where sik represents the arc-length coordinate along the beam centerline at the deformed configuration and is written as dsik = \u221a( \u2202rik \u2202xi )T ( \u2202rik \u2202xi ) dxi = \u2223\u2223 \u2223\u2223 \u2202rik \u2202xi \u2223\u2223 \u2223\u2223dxi. (19) If deformation of a cross section is small, the tangent frame can be used to describe the orientation of the cross section at the constraint definition point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003354_j.automatica.2011.08.004-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003354_j.automatica.2011.08.004-Figure1-1.png", + "caption": "Fig. 1. Sensor kinematics: coordinate axis.", + "texts": [ + " Conclusion and future work are given in Section 5. Solving the sensor kinematics problem is important to determine the linear and angular absolute position of the robot with respect to themeasurements from themice. The kinematic equations obtained in this section will be utilized in Section 3 to determine the best location for N optical mouse sensors. To facilitate understanding, the sensor kinematics problem for only onemouse sensor is first discussed, followed by generalization to the N mice case. Consider the mobile robot shown in Fig. 1, where three coordinate frames are considered to describe the robot motion. In particular, the subscripts A, R and S are utilized to refer, respectively, to the absolute frame placed at the fixed point OA, the robot frame placed at the geometric centerOR of the robot, and the sensor frame placed at a generic point OS on the robot platform. In the following, we will assume that the XR axis is always aligned with the axis of the wheels, so that the YR axis always points toward the forward direction of robotmotion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003288_1.4769758-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003288_1.4769758-Figure1-1.png", + "caption": "FIG. 1. Parallel cylinders of equal radius R separated by a center-to-center distance d floating at an interface separating fluids A and B.", + "texts": [ + " V, we extend our asymptotic approach to the case of a floating cylinder on an arbitrary curved interface. The case of capillary attraction between two cylinders on a curved interface is given in Sec. VI. Section VII summarizes our results and discusses the implications of the results obtained in the present paper to related problems. Consider a static configuration of two horizontal cylinders of infinite length and radius R, density \u03c1s floating on an interface between two fluids of densities \u03c1A and \u03c1B with \u03c1B > \u03c1A, and separated by a center-to-center distance, d, as shown in Figure 1. Our formulation is similar to that given in This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to IP: 142.157.129.89 On: Thu, 27 Nov 2014 17:06:10 Gifford and Scriven.30 Later in the paper, this configuration will be suitably modified to account for background curvature effects and for the case of an array of cylinders. The contact angle, \u03b1 \u2208 [0, \u03c0 ], measured in the lower fluid is the same on either side of each cylinder and is taken to be fixed51 at its equilibrium value, but the position of the contact line is allowed to vary with varying separation distance between the particles", + " For an isolated floating body, the net horizontal force then This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to IP: 142.157.129.89 On: Thu, 27 Nov 2014 17:06:10 vanishes. But in the presence of another floating body, the buoyancy and curvature forces result in an unbalanced horizontal force given by F = ex \u00b7 \u222b c+ b pndl. (5) The above integrals can be evaluated in terms of the parameters shown in Figure 1. In order to present certain geometric expressions in an economical way, we non-dimensionalize all length scales by the radius of the cylinder: non-dimensional variables are indicated with an asterisk and all forces are per unit axial length of the cylinder and are scaled by the surface tension. The relative contribution of buoyancy and curvature forces can be expressed in terms of the Bond number B = g(\u03c1B \u2212 \u03c1A)R2 \u03c3 = R2 l2 c , (6) where lc = \u221a \u03c3/(\u03c1B \u2212 \u03c1A)g is the capillary length. In terms of the parameters shown in Figure 1, vertical equilibrium for each cylinder is expressed by the following dimensionless equation: B { \u03b81 + \u03b82 + 1 2 [sin(2\u03b81) + sin(2\u03b82)] + 2H\u2217 [sin(\u03b81) + sin(\u03b82)] } = 2 [sin(\u03b1 + \u03b81) + sin(\u03b1 + \u03b82)] + 2\u03c0 B D, (7) where D = (\u03c1s \u2212 \u03c1A)/(\u03c1B \u2212 \u03c1A) is the density parameter, and H* = H/R is the non-dimensional height to which the cylinders are displaced vertically relative to a flat interface far from the cylinders. In Eq. (7), the term on the left hand side and first and second terms of the right hand side represent contributions from the buoyancy force, surface tension force, and the weight of the particle, respectively", + " In this limit, the slope of the interface is small, allowing linearization of the curvature about a flat undisturbed interface. Equation (9) becomes d2h\u2217 dx\u22172 = Bh\u2217. (10) The general solution of (10) is h\u2217(x\u2217) = ae\u2212B1/2x\u2217 + beB1/2x\u2217 , (11) This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to IP: 142.157.129.89 On: Thu, 27 Nov 2014 17:06:10 where a, b are constants of integration to be determined by applying boundary conditions. The locations of the contact lines (x\u2217 1 and x\u2217 2 in Figure 1) are geometrically related to the separation distance d. Therefore, the force of attraction depends on four non-dimensional parameters - B, D, \u03b1 and d*, where d* = d/R. From Eqs. (7) and (8), for every solution given by values (D, \u03b81, \u03b82, \u03b1), a complimentary solution (1 \u2212 D, \u03c0 \u2212 \u03b81, \u03c0 \u2212 \u03b82, \u03c0 \u2212 \u03b1) exists which is equivalent to flipping Figure 1 upside down. The basic approach in this paper is to assume a configuration and first equilibrate the vertical force on a particle. We then calculate the shape of the interface due to all particles using Eq. (11) and the associated boundary conditions. Once H*, \u03b81, and \u03b82 are known, the horizontal force of attraction is then calculated. In this section, we obtain the force of attraction for two identical parallel cylinders on a horizontal interface for the geometry shown in Figure 1. This configuration has a symmetry plane midway between the two cylinders. In the present calculations, we calculate the force of attraction of the left cylinder, S1. For convenience, the origin is chosen at the intersection of the undisturbed flat interface and the plane of symmetry. The boundary conditions for the interface between cylinders S1 and S2 are dh\u2217 dx\u2217 = 0, at x\u2217 = 0, (12) dh\u2217 dx\u2217 = \u2212 tan(\u03c61), at x\u2217 = \u2212x\u2217 1 . (13) The interface shape between the two cylinders is found to be h\u2217(x\u2217) = B\u22121/2 tan(\u03c61) [ cosh ( B1/2x\u2217) sinh ( B1/2x\u2217 1 ) ] , \u2200 \u2212 x\u2217 1 \u2264 x\u2217 \u2264 x\u2217 1 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002730_1.4857875-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002730_1.4857875-Figure1-1.png", + "caption": "FIG. 1. Schematic illustration of the process for forming silicon wrinkles on pre-stressed elastomers.", + "texts": [ + "17,18 We developed an approach called rolling-transfer technique, which is more convenient and controllable compared to traditional transfer-printing method,12,23,24 to form wrinkled Si nanomembranes on elastomeric substrate. Furthermore, taking advantage of the fact that Raman spectroscopy is able to probe the strain distribution along depth by using excitation with different wavelengths, the strain and its 3-D distribution in the wrinkled Si nanomembranes has been comprehensively investigated, and a relevant theoretical model has been proposed. The detailed process for rolling-transfer technique is illustrated in Fig. 1. First, a (100) silicon-on-insulator (SOI) wafer with a 30 nm top Si nanomembrane and 340 nm buried oxide (BOX) layer was patterned into rectangles (30 120 lm2), as shown in Fig. 1(a). The patterned SOI wafer was then dipped into HF solution to release top Si nanomembranes [Fig. 1(b)]. Facilitated by van der Waals forces,25 the free-standing nanomembranes were supported by, but not bonded to the underlying substrate without altering the pattern pre-defined [Fig. 2(a)]. Then, the cured polydimethylsiloxane (PDMS) stamp was bonded on a rigid steel cylinder and put into conformal contact with the free-standing Si nanomembranes, as shown in Fig. 1(b). As the rolling of the steel cylinder proceeded, the silicon nanomembranes were detached from the Si substrate and transferred on to the pre-stressed PDMS surface [Fig. 1(c)]. Finally, PDMS was removed from the cylinder and restored into the flat geometry, thus creating the wrinkle-shaped single-crystal Si nanomembranes [Fig. 1(d)]. Superior to traditional transfer-printing method,12,23,24 the proposed rolling-transfer technique can preclude the intractable bubble formation between PDMS and Si nanomembranes, during the conformal contact process, therefore, the transfer efficiency can be significantly enhanced up to 100%. In addition, the geometry including the amplitude and the wavelength of wrinkle-shaped single-crystal Si nanomembranes can be precisely tuned by the radius of the steel cylinder (not shown). The optical microscope image of wrinkled Si a)Authors to whom correspondence should be addressed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003449_imece2011-63452-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003449_imece2011-63452-Figure6-1.png", + "caption": "FIGURE 6: Angle between ring-planet mesh and impulse force vector.", + "texts": [ + " In reality, Fp will vary with p due to formation of loaded-zone and rotation of fault with respect to carrier. In this case, we get a pair of sidebands separated by frequency 2\u03c9pc. The reason for this behaviour is the variation in the angle (\u03b3\u2217) between ring-planet mesh (or line-of-action) and impulse force vector as shown in 8 Copyright c\u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use the fig. 6. The direction of impulse force vector is along the fault located on outer-race, i.e., at an angle \u03b3p in rotating CS. So, when force vector is parallel to ring-planet mesh (\u03b3\u2217 = 0), maximum force is transmitted to ring gear. But when force vector is perpendicular to ring-planet mesh (\u03b3\u2217 = \u03c0 2 ), force transmission is minimum. Therefore, as the fault rotates relative to carrier, mesh-force varies sinusoidally with frequency \u03c9pc. Hence, a pair of sidebands are formed at frequencies p\u03c9d \u00b1\u03c9pc. Notice that for any value if p, the central peak at fault frequency (p\u03c9d) will not be present in the response spectrum" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001190_icelmach.2010.5608143-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001190_icelmach.2010.5608143-Figure3-1.png", + "caption": "Fig. 3 Distributions of (a) magnetomotive force, and (b) permeance.", + "texts": [ + " The operating principle and the transmission torque under the synchronous operation in accordance with the gear ratio are formulated [6]-[12]. High-order components contained in the cogging torque are computed by employing the 3-D FEM and the result of the analysis is verified by carrying out measurements on a prototype. Furthermore, a method for reducing the cogging torque is discussed. II. OPERATING PRINCIPLE Assuming that a low-speed rotor is removed, only a highspeed rotor magnet generates magnetomotive force shown in Fig. 3(a), and stationary pole pieces generate permeance shown in Fig. 3(b), where \u03b8 represents rotor angle. In this model, Fourier series expansions of F(\u03b8) and R(\u03b8) are shown in (1) and (2), respectively. N. Niguchi, K. Hirata, M. Muramatsu, and Yuichi Hayakawa are with the Department of Adaptive Machine Systems, Graduate School of Engineering, Osaka University, Yamadaoka, Suita, Osaka, 565-0871 Japan (e-mail:noboru.niguchi@ams.eng.osaka-u.ac.jp). { }\u2211 \u221e = \u2212= 1 )12(sin)( m hm NmaF \u03b8\u03b8 (1) { }\u2211 \u221e = \u2212+= 1 )12(sin)( l slo NlaRR \u03b8\u03b8 (2) where Nh is the number of pole pairs in the high-speed rotor, Ns is the number of stationary pole pieces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.16-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.16-1.png", + "caption": "Fig. 2.16 The two pallets and the escape wheel in the Grasshopper escapement", + "texts": [ + " Although the grasshopper escapement was not used in his final Harrison Number 4 (H4), which was a watch, it left a mark in history (Fig. 2.14). The Grasshopper escapement was also evolved from the anchor escapement. It has been suggested that the name of this escapement comes from the resemblance of the pallet arms to the legs of a grasshopper. As shown in Fig. 2.15, the escapement consists of an escape wheel, a pendulum, a driving mechanism (the lifted weight) and two pallets shaped like a grasshopper. Figure 2.16 gives the details of the escapement: The right pallet has an elbow joint connected to a heavy tail and a forearm, as well as a composer. The tail is slightly heavier so that the forearm tends to move away from the escape wheel. The composer prevents the forearm from rising further. As the upper arm rotates clockwise, the tip of the pallet at the forearm is pushed downwards. When lifted by the escape wheel, the pallet will take the composer with it, and when released, it will return to the resting position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003583_iros.2011.6094491-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003583_iros.2011.6094491-Figure3-1.png", + "caption": "Fig. 3. Reactive stepping of a humanoid after a push from behind (top row) and its rimless wheel model (botton row). TG is the CoM location and p is the CoP location of the robot. The two legs (spokes) are depicted by solid blue and dotted green respectively. The dotted red arrow indicates the velocity of the CoM. (a) The humanoid is subjected to a push from behind. (b) The humanoid takes a step on the second slope. (c) A follow-up step is executed to bring the robot to a vertically upright position.", + "texts": [ + " GENERA LIZED FOOT PLACEMENT ESTIMATOR ON NON-LEV EL GROUND Our reactive stepping controller is based on a simplified dynamic model of the humanoid: a rimless wheel with only two spokes (See Fig. 2). We assume that the rimless wheel model has massless spokes and a point mass at the center of the wheel. The central idea behind the controller is as follows: when the high level controller decides to execute a step due to a large push, the reactive stepping controller computes the stepping location where the rimless wheel model would come to a complete stop with its CoM directly above the step location as shown in Fig. 3(c). We call this stepping point the generalized foot placement estimator (GFPE). This is an extension of the FPE [7]. Note that we first compute the GFPE and the corresponding configuration of the rimless wheel will be created with its leg angle a defined as the half of the angle between the two spokes (See Fig. 6), which is computed according to the GFPE. Since the rimless wheel model is defined in 2D and the robot exists in 3D, we determine a plane on which the simplified rimless model resides" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001558_iros.2010.5654396-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001558_iros.2010.5654396-Figure1-1.png", + "caption": "Fig. 1. The PVTOL aircraft in presence of crosswind.", + "texts": [ + "00 \u00a92010 IEEE 1606 The PVTOL aircraft is an underactuated system, since it possesses two inputs, u1, u2, and three degrees of freedom, (x,y,\u03c6), and that it moves on a plane. The PVTOL aircraft is composed of two independent motors which produce a force and a moment on the vehicle. The main thrust, u1, is the sum of each motor thrust. The roll moment, u2, is obtained by the difference of motors angular velocities. In real conditions, the aircraft is generally exposed to crosswind. If the PVTOL is affected by a crosswind, the aircraft will be pushed over or rolled away from the wind. Consequently, this leads to include additional forces acting over each rotor, see Figure 1. These forces are due to the airflow generated by the lateral wind. It means that, the magnitude of these forces, is a function of the incoming lateral airflow coming from the wind, see Figure 2. The induced wind speed in a propeller is defined as V = ( f 2\u03c1A ) 1 2 , where f is the thrust generated by the propeller, \u03c1 is the air density and A is the propeller area [14], [22]. The thrust, fkT = fk + fwk , k = 1,2, could be expressed as (see Figure 2) fkT = 2\u03c1AV\u0302Vp (1) where Vp is the induced wind speed in the propeller and V\u0302 is the total induced wind speed by the rotor and lateral wind", + " Moreover, V\u0302 = [ (Vw cos\u03b1 +Vp) 2 +(Vw sin\u03b1)2 ] 1 2 , where \u03b1 is the angle between the rotor axis and the lateral wind axis, with \u03c6 = 0\u25e6 and a wind coming from the right in the x-axis, \u03b1 = 90\u25e6, see Figure 2. It is important to notice that, without lateral wind, Vw = 0, then this gives V\u0302 = Vp, fwk = 0, and (1) becomes fkT = fk = 2\u03c1AV 2 p ; \u2200 k = 1,2, which represents the classical equation of induced wind speed in a propeller. The dynamical model of the PVTOL aircraft, in presence of crosswind, can be obtained from Figure 1 and using Newton - Euler\u2019s approach, mx\u0308 = \u2212( f1T + f2T )sin(\u03c6)+ \u03b5( f1T \u2212 f2T )l cos(\u03c6) my\u0308 = ( f1T + f2T )cos(\u03c6)+ \u03b5( f1T \u2212 f2T )l sin(\u03c6)\u2212mg I\u03c6\u0308 = ( f1T \u2212 f2T )l where x,y denote the horizontal and the vertical position of the aircraft\u2019s center of mass, \u03c6 is the roll angle of the aircraft made with the horizon, m is the total mass of the aircraft, g is the gravitational acceleration, l is the distance between the rotor and the aircraft\u2019s center of mass and I is the moment of inertia. f1T = f1 + fw1 and f2T = f2 + fw2 are the total forces produced by the thrust of the motors f1 and f2 and the forces due to the wind, fw1 and fw2 , in each motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001374_s12206-009-0101-5-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001374_s12206-009-0101-5-Figure2-1.png", + "caption": "Fig. 2. Three-dimensional half model for FEM analysis of induction heating.", + "texts": [ + " Subsequently, a heat flow analysis through the steel plate is then performed to obtain the temperature distribution of the plate, which is used to determine the material properties for the following analysis of the electromagnetic field to obtain the heat flux distribution. This procedure continues until convergence to a solution within an allowable range. To perform the electro-magnetic and heat flow analysis to obtain the heat-flux and temperature distributions of a steel plate with FEM, a solution domain consisting of the steel plate, air and inductor is modeled as shown in Fig. 2. Meanwhile, as the inductor travels on the steel plate, the position of the inductor in the domain changes constantly. A part of air element changes to the inductor, whereas the previous inductor becomes air. To handle this constant change, a quasi-stationary state to the traveling direction of the inductor is assumed for the heat flux and heat flow analyses. Then, the heat-flux generation and heat flow can be modeled with a moving coordinate system, which has the origin at the position of the inductor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000423_j.mechmachtheory.2008.04.005-Figure8-1.png", + "caption": "Fig. 8. (a) A double chain linkage with five non-intersecting pairs. (b) The edge vectors and the inclination angles.", + "texts": [ + " This proof can be easily extended to double chain with any even number of non-intersecting pairs. Although the expressions of Eqs. (21) and (22) looked similar to what we obtained for the double chain with intersecting pairs, they are in fact different. This is due to the way the pieces were arranged under the parallelogram condition. Here p1, p2, p3 and p4 represent the lengths of opposite pieces, see Fig. 7. So do q1, q2, q3 and q4. Now insert an additional pair into the assembly to make a double chain consisting of five non-intersecting pairs, see Fig. 8a. Note that, unlike the previous double chain with four non-intersecting pairs, the pair on the left consists of a piece with edge lengths of p5 and q1 whereas the other piece has edge lengths of q5 and p1 in order to preserve the loop parallelogram condition. The angles sustained by p5 and q1, and by q5 and p1, are represented by a5 and b5, respectively. Plotting all of the vectors p\u2019s and then q\u2019s, under the parallelogram condition, see Fig. 8b, we obtain a1 \u00fe a2 \u00fe a3 \u00fe a4 \u00fe a5 \u00fe h \u00bc 2p \u00f023\u00de and b1 \u00fe b2 \u00fe b3 \u00fe b4 \u00fe b5 h \u00bc 2p: \u00f024\u00de in which h is the motion angle between vectors p1 and q1. Since a\u2019s and b\u2019s are constants, obviously Eqs. (23) and (24) cannot be maintained unless h were constant. This simply shows that the assembly is not mobile. The same is true to all of the double chains with odd number of non-intersecting pairs. In summary, the mobility conditions for a double chain linkage with non- intersecting pairs are (a) The number of pairs much be even, and (b) The vector sum of edges of the pieces in every other pairs must be zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000092_iet-cta:20060370-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000092_iet-cta:20060370-Figure1-1.png", + "caption": "Fig. 1 Frame definition and geometry of the engagement", + "texts": [ + " Second, the Lyapunov stability criterion [17] is utilised to study the stability properties of the resulting FCS, as well as to determine the relation between the asymptotic stability and the output of the fuzzy inference system (FIS). Third, the performance of the derived FCS is studied and compared with other control schemes in a realistic engagement, in which case the PID gains of all the considered processes are tuned by means of the genetic algorithm [18], whose efficiency and effectiveness has been recognised in tuning of the fuzzy modules and PID controllers [6, 7, 19, 20]. As illustrated in Fig. 1, the proposed engagement is a surface-to-air tactical missile interception scenario. Without loss of generality, the thrust P, gravitational force G and atmospheric force R will be considered throughout the engagement. In the present case, the thrust acts in the direction of the axis xbo of the body frame, the gravitational force acts in the opposite direction of the axis OY1 of the inertial frame, and in the velocity frame (throughout this paper, the word \u2018velocity\u2019 will only be used to designate a vector quantity: the corresponding scalar will be called the speed) the atmospheric force can be expressed simply IET Control Theory Appl" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002268_j.surfcoat.2010.01.010-Figure11-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002268_j.surfcoat.2010.01.010-Figure11-1.png", + "caption": "Fig. 11. (a) An isothermal section at 950 \u00b0C in the Nb-Ti\u2013Fe ternary alloy system, (b) the liqui laser power of 1.1 kW and \u03c4=0.024 s (within ESI time range) and (d) suggested solidificatio time range) [12].", + "texts": [ + " 10) matches only the standard diffraction angles of Fe2Nb and Fe2Ti (or \u03b5) compounds. This is ion angles of the identified phases; P=1.1 kW; (a) \u03c4=0.024 s. consistent with Table 2 which shows an increasing Fe content (i.e. increasing substrate dilution) with an increase in interaction time. At \u03c4=0.08 s, this causes the clad microstructure to change significantly from Ti and Nb-rich phases to Fe-rich phases. The above experimental results can be interpreted with the aid of the ternary Nb\u2013Ti\u2013Fe phase diagram presented in Fig. 11. Fig. 11(a) shows an isothermal section of the Nb\u2013Ti\u2013Fe ternary alloy system at 950 \u00b0C (i.e. a temperature at which no liquid is present) while (b), (c) and (d) illustrate the liquidus surface projection in this system which can be used to predict the solidification path for liquid alloys being cooled at higher rates than required to achieve equilibrium solidification. Superimposed on all these diagrams are the bulk clad compositions reported in Table 2 for \u03c4=0.024 and 0.034 s represent- ing ESI and SI time ranges respectively. (Note that due to their similar compositions, \u03c4=0.027 s lies close to the 0.024 s data point in Fig. 11 and \u03c4=0.030 or 0.040 s lies close to the 0.034 s data point.) Fig. 11(c) and (d) also show solidification paths which will be discussed in more detail below. Under equilibrium solidification conditions, and assuming that the laser cladding molten pool obtained a uniform bulk composition throughout the clad section before solidification starts, Fig. 11(a) predicts an equilibriummicrostructure consisting of \u03b2-phase and FeTi compound at 950 \u00b0C for \u03c4=0.024 s. The XRD results are consistent with this prediction except for the additional presence of Ti4Nb. The microstructural features indentified using EDS analysis indicate a matrix which is Fe-rich and consistent with the FeTi compound predicted by Fig. 11(a), but with two Ti\u2013Nb solid solution phases (one Ti-rich, and one Nb-rich) rather than a single Ti\u2013Nb rich \u03b2-phase predicted by Fig. 11(a). The clad bulk composition for \u03c4=0.034 s ard diffraction angles of identified phases; P=1.1 kW; \u03c4=0.080 s. located on Fig. 11(a) indicates that, under equilibrium solidification, the clad microstructure should consist of the \u03b2 and FeTi phases as before but an additional Fe-rich intermetallic phase, \u03b5 (both Fe2Nb or Fe2Ti). All three of these phases are confirmed by the XRD pattern illustrated in Fig. 9. However, the results of XRD, SEM and EDS analyses also show the presence of a ternary solid solution FeNb(Ti) intermetallic phase, \u03bc, in the matrix. Finally, for \u03c4=0.080 s, there is no microstructural analysis available but the XRD pattern which is revealing only strong peaks of both Fe2Nb and Fe2Ti (\u03b5 phase). It is argued that based on increasing dilution from the substrate melting and the concomitant increasing Fe content in the clad melt pool by increasing the interaction time, the clad bulk composition would lay within \u03b5 region in Fig. 11(a). The discrepancies between the experimental observations for the ESI and SI time ranges and the equilibrium predictions of Fig. 11(a) can be understood by considering the very high rates of the laser dus surface projection, (c) suggested solidification path for the clad melt pool made at a n path for the clad melt pool made at a laser power of 1.1 kW and \u03c4=0.034 s (within SI cladding process. The short interaction times encountered under the ESI and SI time ranges create a non-uniform clad melt pool bulk composition and high solidification rates, which are far from equilibrium conditions. The influence this has on phase formation will be discussed in detail below", + " In addition, since not all Nb powder particles are dissolved into the molten pool, the liquid will have a Ti-rich composition compared to the bulk clad composition and the intended Ti\u201345 wt.% Nb ratio. At \u03c4>0.022 s, the microstructural evidence indicates that all of the Nb particles have melted but there is insufficient time to create a uniform liquid composition such that when solidification begins there remains Ti-rich and Nb-rich molten regions. For ESI time range where dilution by the Fe substrate is low this creates liquid compositions close to those plotted in Fig. 11(c) labeled \u03b2Nb and \u03b2Ti. According to the Ti\u2013Nb phase diagram (Fig. 13), and considering the chemical composition of the coarse dendrite phase either in Fig. 6 and Fig. 4(b) (Spectrum 3 either in Table 3 or Table 4), it can be stated that the coarse dendrite phase is \u03b2Nb (\u223cTi\u201359wt.%Nb excluding the Fe content, based on Table 3) containing a small weight fraction of Fe atoms dissolved in solid solution. This is consistent with Poppov et al. [9] who have reported that the \u03b2-phase in the Ti\u2013Nb system has a BCC structure and an increasing lattice parameter with the Nb content reaching 3", + " Moreover, lattice parameters of the orthorhombic \u03b1\u2033 which are dependent upon its Nb content, are reported as a0=3.099 \u00c5, b0=4.893 \u00c5 and c0=4.641 \u00c5 for the \u03b1\u2033 phase based on Ti-30wt.%Nb [10]. The XRD diffraction pattern shown in Fig. 8 confirms the presence of an orthorhombic crystal structure Ti4Nb which corresponds to the \u03b1\u2033 phase. The lattice parameter for the Ti4Nb ) Ti\u2013Nb phase diagram [11]. measured from the XRD (i.e. a0=3.166 \u00c5, b0=4.854 \u00c5 and c0= 4.652 \u00c5 according to Table 6) are very close to those reported in the literature. Fig. 11(c) shows the predicted solidification path for a cladmade in the ESI time range. Superimposed on the liquidus surface projection of Nb\u2013Ti\u2013Fe system is the clad bulk composition at \u03c4=0.024 s as well as the the Nb-rich \u03b2 (\u03b2Nb) and the Ti-rich \u03b2 (\u03b2Ti) approximate compositions reported in Table 4 (i.e. Spectrum 3 and Spectrum 2, respectively). These two compositions are taken as representative for the Nb-rich and Ti-rich molten regions throughout the clad melt pool before solidification begins. The solidification paths shown in Fig. 11(c) are more specifically illustrated in Table 7. For ESI time range, primary dendrites of \u03b2Nb are the first solid to form (i.e. stage 1a), followed by solidification of primary dendrites of \u03b2Ti at lower temperatures (i.e. stage 1b). (Note: Following solidification the \u03b2Ti will transform martensitically to \u03b1\u2033 via fast cooling in solid state as mentioned above). Simultaneous freezing of secondary dendrites of \u03b2Ti and FeTi begins at the \u03b2/FeTi phase boundary (i.e. stage 2, depicted as the path u2\u2013e1 in Fig. 11(c)). Due to the high difference in the melting point of Nb with that of either Ti or Fe, the liquidus surface is steeply sloped downward from the Nb rich corner to either of the Ti or Fe portions. Considering also the XRD results for ESI time range which do not reveal the presence of \u03b5, it is argued that the solidification path does not include the ternary eutectic point u2. The final solidification event in the alloy is the freezing of the interdendritic liquid through the binary eutectic reaction at point e1 to form \u03b2Ti and FeTi (i", + " 4(b) and the XRD results indicating the presence of FeTi confirms this final sequence in the solidification path. The primary difference between the laser cladding at ESI and SI time ranges is an increased dilution of the Fe substrate which increases the Fe content of the molten pool. It is expected that the interaction times are still low enough that Nb-rich and Ti-rich liquid regions still exists under these conditions. Accordingly, the solidification path predictions for clad sections made under the SI time range are given in Fig. 11(d), with the detailed path description indicated in Table 7. Owing to the elevated Fe content of the Nb-rich liquid at \u03c4=0.034 s, the first solidification event is the formation of primary dendrites of \u03bc phase (i.e. stage 1a). This explains the presence of \u03bc phase in the XRD pattern of Fig. 9. This is followed by solidification of primary fine dendrites of \u03b2Ti at lower temperatures (i.e. stage 1b) in Ti rich regions of the liquid. The next stage, stage 2, is the simultaneous freezing of secondary dendrites of \u03b2Nb and \u03b5 through \u03b2/\u03b5 phase boundary (i.e. the path u1\u2013u2 in Fig. 11(d)). This explains the presence of \u03b5 phase found in the spectrum of Fig. 9. The final solidification event is the freezing of the interdendritic liquid through the ternary eutectic reaction point u2 through which \u03b2Ti, \u03b5 and FeTi form in a lamellar structure. For \u03c4=0.08 s, the dilution from the substrate is obviously higher and the clad bulk chemical compositionmay fall within the \u03b5 region in the isothermal section of the Nb\u2013Ti\u2013Fe system (shown in Fig. 11(a)). Therefore, considering the higher possibility of obtaining a uniform molten pool composition in LI time range, a final clad microstructure containing only \u03b5 (i.e. a mixture of Fe2Nb and Fe2Ti) is expected which is in agreement with the XRD results listed in Table 6. Utilizing the highly concentrated power of the fiber laser along with high scanning speeds, crack and pore- free coatings with excellent hardness (up to \u223c1000 HV0.05) were deposited by preplaced laser cladding of a premixed powder of Ti-45wt" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001114_0951192x.2010.528033-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001114_0951192x.2010.528033-Figure14-1.png", + "caption": "Figure 14 shows the manufacturer and the estimated WS. A simple comparison between the two figures illustrates the resemblance in the WS. Two sections are taken at A and B as shown to demonstrate the similarity between the manufacturer and the obtained WS. In this figure, the Unimate 9000 robot WS as provided by the manufacturer is shown on the left side. While, the estimated WS produced by the developed algorithm is shown on the right side. Two slicing planes are taken at different Z positions to illustrate the similarity between the two WSs. Manufacturers could provide 3D models of WS/SV in free space. Hence, the power of the developed algorithm is illustrated here by taking a slicing plane at different Z position to show the resemblance between the two. Also, points accessible by the robot are automatically", + "texts": [ + "0000 0.1500 0.0000 72.0071 2.0071 3 R q3 0.0000 0.0000 3.1415 73.9269 3.9269 4 P 0.0000 q4 0.1000 0.0000 0.1500 0.2500 PUMA 760 (Figure 13) 6 0 \u2013 0.0000 0.0000 0.0000 0.0000 \u2013 \u2013 1 R q1 1.0300 0.0000 1.5708 71.2217 4.3633 2 R q2 0.0000 0.6500 0.0000 70.3491 3.4907 3 R q3 0.0000 0.0000 1.5708 70.7854 3.9270 4 R q4 0.6000 0.0000 71.5708 74.6426 4.6426 5 R q5 0.0000 0.0000 1.5708 71.9199 1.9199 6 R q6 0.1000 0.0000 1.5708 74.6426 4.6426 Note: All lengths are in metres and all angles are in radians. Figure 14. WS of Unimate 9000: (a) manufacturer WS, (b) estimated WS. D ow nl oa de d by [ FU B er lin ] at 0 5: 20 1 4 M ay 2 01 5 genetic search parameters are (Goldberg 1989) as follows: The search results (optimum base placement and corresponding OF value) are given for the six robots that have feasible solutions in Table 3. This table provides the coordinates of each robot within the workplace and its orientation with respect to the reference point identified in the work cell. Convergence of genetic search is achieved when the average OF for the whole population becomes very close to that of the best member in the population (Goldberg 1989)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000817_s10845-009-0346-y-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000817_s10845-009-0346-y-Figure5-1.png", + "caption": "Fig. 5 Accelerometer location", + "texts": [ + " The input spur gear is a 40-tooth gear, driving three, input spur idler gears of 72 teeth. The idler shafts drives three, 48- tooth output spur idler which drives a single 64-tooth output spur gear. Accelerometers where mounted on the input drive pinion and on each output drive idler. The fault was characterized as removal of 20% of the gear tooth on one of the output drive idler. The location of each accelerometer was placed such that: there was optimal fault detection, and that there would be some potential for sensor fusion (Fig. 5). The STG test rig is representative of an operational STG in the generation of synchronous tones test CBM/HUMS analysis gear analysis algorithm. The cost, size and power where driving considerations in the design of this STG, which was ultimately developed by SpecrtaQuest, Inc. A more representative gearbox design, such at seen in Krantz (1996) from the Comanche STG is given in Fig. 3. Among the traditional technique for detection of the localized gear faults, the time domain synchronous average (TSA) (Combet and Gelman 2007) is the most popular technique", + " Theoretically, this would improve the signal to noise of the gear under analysis, which has the benefit of increasing the discriminate capability of the gear algorithms. Testing of the beam former showed no improvement gear fault discrimination. This was disappointing and, on further investigation, not surprising. In RADAR for example, each antenna element is a dipole, with sensitivity not a function of arrival angle. This is not the case of shear style accelerometer, which has sensitivity of, at most, 10\u25e6 on axis. Due to the placement of the accelerometers (see Fig. 5) there was very little signal interference from the adjacent gears. Due to this limitation, additional accelerometer mounting bosses where machined into the gearbox shaft supports. This will allow future analysis of beam former techniques in the future. Narrowband interferance canceller Most gear algorithms used in HUMS are based on measuring frequencies associated with gear mesh tones. For example, Mesh Analysis is the ratio of gear mesh harmonics to the base mesh tone. Similarly, G2 analysis is the ratio of the signal average peak to peak value and the gear mesh tone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002372_titb.2012.2226596-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002372_titb.2012.2226596-Figure1-1.png", + "caption": "Fig. 1. (a) Major muscle groups used in the musculoskeletal model. (b) Robotic orthosis leg model to guide subject\u2019s limbs on reference trajectories (side view).", + "texts": [ + "00 \u00a9 2012 IEEE obtained from this simulation is important for proper elucidation of clinical studies involving robotic gait training orthoses, and to develop a better understanding of the effects of cadence regulation on muscle function. A 2-D musculoskeletal model was developed for simulation purpose. The body segments included in the model were trunk, thigh, shank, and foot. Trunk was included as a single rigid body to account for the upper extremity inertial parameters. The segments were connected with revolute joints and the gait was considered as purely sagittal [see Fig. 1(a)]. The model was developed for a subject 1.78-m tall and weighing 77.9 Kg according to the segment lengths and inertial parameters reported in [39]. The dynamical equation of motion [32] was derived by Newton\u2013Euler Principle and is given by M1 (\u03b8) \u03b8\u0308 + C1 ( \u03b8, \u03b8\u0307 ) + G1 (\u03b8) + R (\u03b8) FM + F ( \u03b8, \u03b8\u0307 ) = 0 (1) where \u03b8, \u03b8\u0307, and \u03b8\u0308 are vectors of generalized position, velocity, and acceleration, respectively. M1(\u03b8) is the system mass matrix, C1 is a vector of centrifugal and Coriolis torques, and G1 is the vector of gravitational torques. FM is a vector of muscle activations, R (\u03b8) is the matrix of muscle moment arms, and F (\u03b8, \u03b8\u0307) is a vector of ground contact force (GCF). Eight major muscle groups of each leg were selected for simulation purpose in this study [see Fig. 1(a)]: soleus (SOL), gastrocnemius (GAS), tibialis anterior (TA), hamstrings (HAM), vasti (VAS), gluteus maximus (GLU), uniarticular hip flexors (iliopsoas, IP), and Rectus Femoris (RF). The muscle attachment parameters were determined according to anatomical data reported in [39]. All the muscles were modeled as mass-less wires. The GCF was modeled as nonlinear viscoelastic spring damper arrangement. The spring damper arrangement was positioned vertically at the heel and toe of each foot [40] and was characterized according to the relation Fz = az3(1 \u2212 bz\u0307) (2) where z and z\u0307 were the vertical deformation and rate of vertical deformation of heel and toe, respectively. The values of a and b were 0.25 \u00d7 109 N/m3and 1.0 s/m, respectively [40]. 1) Model: The purpose of a robotic orthosis in this study was to guide subject\u2019s limbs on reference trajectories. Each leg of the robotic orthosis weighing 9.8 Kg was modeled as a double pendulum [41] with an additional shoe section [see Fig. 1(b)]. The robotic orthosis has revolute joints at hip, knee, and ankle to provide motion in sagittal plane. The orthosis joints were modeled to be in perfect alignment with musculoskeletal model joints. The dynamics of robotic orthosis is given by M2 (\u03b8) \u03b8\u0308 + C2 ( \u03b8, \u03b8\u0307 ) + G2 (\u03b8) = Trob (3) where M2(\u03b8) is the orthosis mass matrix, C2 is a vector of orthosis centrifugal and Coriolis torques; G2 is a vector of orthosis gravitational torques. Trob is the vector of torque applied by robotic orthosis", + " The mass matrices, centrifugal, Coriolis, and gravitational torques from musculoskeletal, and orthosis model ((1), (3)) were combined to form an integrated dynamic model of the system and is represented by M (\u03b8) \u03b8\u0308 + C ( \u03b8, \u03b8\u0307 ) + G (\u03b8) + F ( \u03b8, \u03b8\u0307 ) \u2212 Trob = R (\u03b8)FM = Tj (4) where M , C, and G are the combined terms for robotic orthosis and subject. The maximum joint ranges of motion were kept under anatomical constraints by the orthosis torques. Tj is the human joint torque vector. The integrated dynamic model in (4) was formed such that the robot applied torques Trobwere subtracted from the other matrices in order to calculate the human joint torque vector Tj . The robotic orthosis torques at hip, knee, and ankle joint was produced by the antagonistic actuation of pneumatic muscle actuators (PMA) [see Fig. 1(b)]. For the purpose of this study, a three element dynamic model of PMA developed by Reynolds et al. in [42] was used and is given by Mix\u0308 + B (P ) x\u0307 + K (P )x = Fce (P ) \u2212 Mg (5) where x, x\u0307, x\u0308, are the amount of PMA contraction, contraction velocity, and acceleration, respectively. Fce is the effective force provided by the contractile element, and K and B are spring and damping coefficients, respectively. All these coefficients are functions of pressure P . Mi is the inertial load and Mg is the load due to the weight of the system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001779_s1674-5264(09)60095-8-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001779_s1674-5264(09)60095-8-Figure2-1.png", + "caption": "Fig. 2 Mathematical model of a vibrating screen", + "texts": [ + " At each point the vibration is a combination of the translation of the center of gravity and the screen pitching about the center of gravity. Previous studies neglected the influence of elastic forces in the horizontal and vertical direction on the swing of the vibrating screen[3,11]. An accurate dynamic model consisting of three differential equations that include coupling of degrees of freedom in the vertical, horizontal and swing directions is proposed. The mathematical model of the vibrating screen is shown in Fig. 2. The center of gravity O, is taken as the origin of a rectangular coordinate system at static equilibrium, in accordance with rigid motion on the plane[12]. Simultaneous differential equations in generalized coordinates using center of gravity coordinates, (x, y), and the swing declination angle, \u03b8 , may be written as 0 0 0 1 2 2 2 2 1 1 2 cos 2 sin 2 sin( ) sin ( ) cos ( ) cos ( ) x x y y x y y Mx A t f x k x My A t f y k y J Al t f k L L x k L L y k L L \u03b8 \u03c9 \u03c9 \u03b8 \u03c9 \u03b2 \u03b8 \u03b1 \u03b1 \u03b1 \u03b8 = \u2212 \u2212 = \u2212 \u2212 =\u2212 \u2212 \u2212 \u2212 + \u2212 \u2212 \u2212 + (1) where M is the mass of the vibrating screen; J the moment of inertia of M relative to the center of gravity, O; x and y the displacements in the x and y directions; x and y the velocities in the x and y directions; x and y the accelerations in the x and y directions; \u03b8 is the swing angular displacement; \u03b1 the installation angle; xf , yf and f\u03b8 the damping coefficients in the x, y and \u03b8 directions; xk and yk the stiffness coefficients of the supporting spring along the x and y directions; 0A the amplitude of the exciting force, given by 2 0A mr\u03c9= , where r is the radius of eccentricity, m the mass of the eccentric block and \u03c9 the exciting angular frequency; 1L and 2L the distances between each supporting spring and the center of gravity; l the distance between the rotating center of the eccentric block and the center of gravity; and, \u03b2 the included angle between the l and x directions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002460_iet-cta.2009.0622-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002460_iet-cta.2009.0622-Figure6-1.png", + "caption": "Fig. 6 Steady-state periodic orbit in (r1,a1,b1) coordinates", + "texts": [ + " Let this minimum period be denoted T. Notice how each vehicle\u2019s behaviour is identical to that of the next, only shifted T/5 seconds in time. More generally ai(t) = a1(t + T (n \u2212 i + 1)/n) (7) for i = 1, 2, . . . , n. The same result holds for the coordinates ri and bi, so similar plots have not been included. Although the complete multivehicle system lives in R3n, this repetition of patterns from one vehicle to the next suggests that it may be possible to reduce the analysis problem to one in R3. For example, Fig. 6 shows the resulting 392 & The Institution of Engineering and Technology 2011 steady-state periodic orbit in the (r, a, b)-space when n = 5 (corresponding to Fig. 4). Although the data used to generate Fig. 6 is for vehicle i = 1, the orbit is identical to the case of i = 2, 3, . . . , 5, merely shifted 1/5th of a cycle in time (not space). When the number of vehicles n is even, it is possible to conjecture the existence of steady-state solutions other than the equilibrium solutions revealed by Theorem 1. It is also possible to draw some conclusions about the shape of the resulting periodic solutions. In this section, we also focus on the formation\u2019s symmetry in space. First, let ji = (ri, ai, bi) and define j = (j1, j2, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001547_iciea.2009.5138555-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001547_iciea.2009.5138555-Figure3-1.png", + "caption": "Fig. 3. Approximate waveforms of stator currents and back-emfs for phase advanced commutation.", + "texts": [ + " If a fast inner control loop is applied to the drive such as the current control loop or the direct torque control loop, the PWM duty cycle is controlled in one switching cycle to achieve demanded control goal. In addition, the inverter gating signal patterns can be properly selected such as the soft chopping patterns [8]. However, because of the inductive stator winding time constant, the actual stator current lags the demanded current as shown in Fig. 2. Hence, a phase advanced commutation scheme is desired to modify the actual current waveform to improve the drive performance as shown in Fig.3. As compared to the BLDC motor, the IPM-BLDC motor drive needs a larger phase advanced angle to result in the stator current with a larger leading angle such that a better performance could be expected. The detailed features will be discussed in the next section. As reference to Fig. 1 is the power stage of a three phase BLDC motor drive. In the BLDC motor model, SR and SL are the equivalent phase winding resistance and decoupled phase equivalent inductance respectively. Three phase back-emfs are denoted as ane , bne and cne respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003830_icmech.2013.6518544-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003830_icmech.2013.6518544-Figure2-1.png", + "caption": "Fig. 2: Coordinate systems . ow and ob stand for the origins of the world and body coordinate frames, respectively. The foot coordinate frames are fixed to the foot soles.", + "texts": [ + " Three joint axes are positioned at the hip, two joints are at the ankle and one at the knee (Fig. 1). The numerical values of the parameters (Table ) are taken to match our experimental humanoid robot SURALP (Sabanci University Robotics Research Laboratory Platform) [16]. The details of contact modeling and simulation algorithm are in [17]. The modeled reaction forces suffer from peaks, so that Kalman filter is used for smoothing the modeled reaction forces. The coordinate frames are shown in Fig. 2. All the measurements and calculation are in the world frame. The transformation is done using the rotational matrix obtained by the author in [18]. The body frame has an offset offsetx , The role of this offset parameter is to place the center of the support polygon exactly below the center of mass of the robot as shown in Fig. 3. According to the figure it has a negative value. Kalman filter parameters are listed in Table II. , ,Q R and the low pass filters constants are chosen by trial and error" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002991_iset-india.2011.6145347-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002991_iset-india.2011.6145347-Figure6-1.png", + "caption": "Fig. 6. 6/4 SRM", + "texts": [ + "74\ud835\udc3c2\ud835\udc4e\ud835\udc5f\ud835\udc5f\ud835\udc4e\ud835\udc66 At maximum power, \ud835\udc51\ud835\udc43\ud835\udc4e\ud835\udc5f\ud835\udc5f\ud835\udc4e\ud835\udc66 \ud835\udc51\ud835\udc3c\ud835\udc4e\ud835\udc5f\ud835\udc5f\ud835\udc4e\ud835\udc66 = 0 Solving these equations, the maximum power \ud835\udc43\ud835\udc5a\ud835\udc4e\ud835\udc65, corresponding array voltage \ud835\udc49\ud835\udc5a\ud835\udc4e\ud835\udc65 and current \ud835\udc3c\ud835\udc5a\ud835\udc4e\ud835\udc65 can be calculated and are given in Table I. SRM has saliency in both stator and rotor. The stator will have concentrated type of winding and the rotor has no winding. Thus rotor is of simple construction. The principle of operation of SRM is based on the fact that a piece of magnetic material tends to align itself in the minimum reluctance position when placed in a magnetic field [6]. Fig. 6 shows cross sectional view of a 6/4 SRM. When any of the phase is excited, the rotor pole disposes itself to align with energized stator phase, thus minimizing the reluctance in the magnetic circuit. Referring to Fig. 6, if phase A is energized no torque is produced because of alignment of rotor pole with energized stator phase. However if phase C is energized, clockwise torque acts on the rotor. If phase B is energized, anticlockwise torque acts on the rotor. To produce continuous rotor motion, phases should be excited in proper sequence and at correct rotor position. The mathematical equations describing the dynamic behavior of SRM are given below [7]: \ud835\udf03\ud835\udc60 = 2\ud835\udf0b ( 1 \ud835\udc41\ud835\udc5f \u2212 1 \ud835\udc41\ud835\udc60 ) (2) \ud835\udf03\ud835\udc65 = ( \ud835\udf0b \ud835\udc41\ud835\udc5f \u2212 \ud835\udefd\ud835\udc5f ) (3) \ud835\udf03\ud835\udc66 = \ud835\udf0b \ud835\udc41\ud835\udc5f (4) \ud835\udc49 = \ud835\udc51\ud835\udf13\ud835\udc56(\ud835\udf03, \ud835\udc3c\ud835\udc56) \ud835\udc51\ud835\udc61 +\ud835\udc45\ud835\udc56 (5) \ud835\udf13(\ud835\udf03, \ud835\udc3c\ud835\udc56) = \ud835\udc3f(\ud835\udf03)\ud835\udc3c\ud835\udc56 (6) \ud835\udc47 = 1 2 \ud835\udc51\ud835\udc3f \ud835\udc51\ud835\udf03 \ud835\udc3c2\ud835\udc56 (7) \ud835\udc51\ud835\udf14 \ud835\udc51\ud835\udc61 = 1 \ud835\udc3d (\ud835\udc47 \u2212 \ud835\udc47\ud835\udc59 \u2212 \ud835\udc4f\ud835\udf14) (8) \ud835\udc51\ud835\udf03 \ud835\udc51\ud835\udc61 = \ud835\udf14 (9) \ud835\udc47\ud835\udc59 = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002316_acc.2011.5991049-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002316_acc.2011.5991049-Figure1-1.png", + "caption": "Fig. 1. Planar surface vessel schematic.", + "texts": [ + " If the resulting nonlinear optimization problem is feasible, the sliding mode controller will avoid these trajectories. If the optimization problem is infeasible, then the system is over specified and it is not possible to avoid the specified state trajectories. In this case, either the constraints on the control actuation, the constraints on the dynamic state trajectory, or both must be relaxed in order to obtain a feasible optimization problem. The 3-DOF planar model of a surface vessel shown in Figure 1 is considered in this work. This model includes surge, sway, and yaw motion with two propeller force inputs f1 and f2. The geometrical relationship between the inertial reference frame and the vessel-based body-fixed frame (located at the vessel center of mass) is defined in terms of velocities as x\u0307 = vx cos \u03b8 \u2212 vy sin \u03b8 (1) y\u0307 = vx sin \u03b8 + vy cos \u03b8 (2) \u03b8\u0307 = \u03c9 (3) where x and y denote the position of the center of mass, \u03b8 is the orientation angle of the vessel in the inertial 978-1-4577-0081-1/11/$26" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000844_tmag.2007.916121-Figure8-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000844_tmag.2007.916121-Figure8-1.png", + "caption": "Fig. 8. Real part of the deformation at 6000 Hz.", + "texts": [], + "surrounding_texts": [ + "The better correlation of the body sound index for 3333 and 6000 Hz can be understood looking at the mode shapes at both frequencies in Figs. 7 and 8. It is clearly seen that the deformation correlates better in terms of shape and magnitude in the 3333-Hz case than in the 6000-Hz case. A closer look at the 3-D solution shows that the 3333-Hz case is an almost pure 2-D mode shape, where the deformation at 6000 Hz varies strongly along the axial direction of the machine. A previous study [4], however, showed that 6000 Hz was one of the most dominant frequency regarding the acoustic noise radiated by this SRM. Therefore, it can be concluded that the particular caution has to be taken when analyzing electrical machines by means of 2-D structure-dynamic simulation, especially if the machine is face-mounted." + ] + }, + { + "image_filename": "designv11_3_0000466_iros.2008.4650802-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000466_iros.2008.4650802-Figure7-1.png", + "caption": "Fig. 7. Rope permutation strategy.", + "texts": [ + " 6(a), the description of the intersection is the following:\u23a7\u23a8\u23a9El \u2212 C {+,\u2212} 1 \u2212 C {+,\u2212} 1 \u2212 Er \u2212\u2192 El \u2212 C {+,\u2212} 1 \u2212 C {+,\u2212} 2 \u2212 C {+,\u2212} 2 \u2212 C {+,\u2212} 1 \u2212 Er In the case shown in Fig. 6(b), the description of the intersection is the following:\u23a7\u23aa\u23aa\u23a8\u23aa\u23aa\u23a9 El \u2212 \u0302 C {+,\u2212} 1 \u2212 \u0302 C {+,\u2212} 1 \u2212 Er \u2212\u2192 El \u2212 \u0302 C {+,\u2212} 1 \u2212 C {+,\u2212} 2 \u2212 C {\u2212,+} 3 \u2212 \u0302 C {+,\u2212} 1 \u2212C {+,\u2212} 2 \u2212 C {\u2212,+} 3 \u2212 Er The difference between the two types of rope permutation depends on the grasp type used in the loop production. Rope permutation strategy The process of rope permutation is shown in Fig. 7. Both fingers are moved while remaining parallel. At some point, two sections of the rope engage each other by virtue of friction. By continuing to move the two fingers in parallel, the two sections of the rope are permuted (exchange places). At the moment of permutation, the grasp force is increased. In order to achieve smooth permutation, it is necessary to increase the distance between both fingers at this moment. Therefore, grasp force control based on tactile feedback is needed [1]. Rope pulling is an operation that basically removes the intersection, as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002524_3.5535-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002524_3.5535-Figure2-1.png", + "caption": "Fig. 2 Possible exoatmospheric motions.", + "texts": [ + " Such cases might prove interesting for future investigations. For the constant roll rate case (p = 0), the result, Eq. (14), reduces to an expression similar to that derived in ReL 1 for the negative precession mode. IV. Initial Re-Entry Conditions Before re-entry, the vehicle is in a state of moment-free motion, and the relation between the various angular rates is uniquely determined. In general, the vehicle will have some initial roll rate po and will undergo a steady precession (coning motion) with a precession rate 0 and cone half-angle a (Fig. 2a). The cone axis will, in general, be inclined to the flight path, and the velocity vector may lie inside the coning circle (Fig. 2b) or outside the coning circle (Fig. 2a). Special limiting cases of exoatmospheric motion, in which the vehicle is at angle of attack with no coning or is coning symmetrically about the flight path, are shown in Figs. 2c and 2d, respectively. The exoatmospheric roll rate, precession rate, and cone half-angle are related by the expression This relation follows from the second of Eqs. (4) with 6 = 0 = co = 0, and a and 12 substituted for 6 and ^, respectively. After entering the atmosphere, the statically stable vehicle is subjected to an aerodynamic pitch moment that tends to align the vehicle with the flight path", + " The residual coning motion, in the same direction as the angular momentum vector (clockwise), has been called nutation by Nicolaides,7 and the retrograde precession in the opposite direction (counterclockwise) has been called, simply, precession.^ In general, the two motions exist simultaneously. For the special case of Fig. 3c, in which the vehicle is coning symmetrically about the flight path in the positive direction (direct precession), the pitch moment tends to induce a precession in the opposite direction. This condition is therefore quasi-stable, but it can persist throughout the trajectory. For the case shown in Fig. 2b, in which the coning is initially asymmetric about the velocity vector, an instability can occur in which the motion changes from direct precession to retrograde precession. This is discussed later. ^ These definitions are more restrictive than are the classical definitions of nutation and precession as being variations in 0 and if/, respectively (see, for example, Ref. 9, p. 432). D ow nl oa de d by N O R T H D A K O T A S T A T E U N IV E R SI T Y o n Ja nu ar y 25 , 2 01 5 | h ttp :// ar c. ai aa " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001743_3.30282-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001743_3.30282-Figure1-1.png", + "caption": "Fig. 1 Coordinate system geometry.", + "texts": [ + " The Ballistic Research Laboratories have developed a yawsonde of different design and have successfully instrumented a number of projectiles to obtain data on yawing and rolling motion. This paper describes the onboard solar aspect sensing instrument, calibration and analysis techniques, data handling and reduction processes (manual and automated), accuracy limitations, results of recent firings, and limitations of and usefulness of this technique. D ow nl oa de d by U N IV E R SI T Y O F M IC H IG A N o n Fe br ua ry 1 9, 2 01 5 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /3 .3 02 82 The solar aspect angle a (Fig. 1) can be measured by a simple geometric arrangement of sensing cells and slits. Two silicon solar cells are mounted in fixtures which define almost planar fields of view (Fig. 2). Each field of view is established by a slit, serrated along its length to absorb internal light reflections, with reflective surfaces at each end of the slit to permit wide viewing in the plane. The angle 2ft between the two fixtures in a plane perpendicular to the projectile axis (Fig. 3b) is set to a predetermined value", + "9 The method currently in use at the BRL is a modified version of the WOBBLE program. Geometric Relationships It is convenient to define a cartesian coordinate system whose origin is fixed at the center of mass (c.m.) of the projectile and moves with the projectile along its trajectory. The X axis of the system is tangent to the trajectory and positive in the direction of increasing arc length; the Y axis is perpendicular to the plane of the trajectory. This \"trajectory\" or \"range\" coordinate system is shown in Fig. 1 relative to an Earth-fixed system. Let um be a unit vector along the axis of the projectile and us be a unit solar vector directed from the c.m. to the sun; us is constant if the sun does not change position during the flight of the projectile and if the apogee altitude is small compared to the distance to the sun. Then coso- = u\u00ab-uw (2) where the vectors are defined in the range coordinate system as (wi,w2,m3) and us = (3) The solution to the equations of motion10 of a symmetrical shell for small amplitude motion is expressed by f = K&** + K#\u00bb* + \u00a30 (4) where |\u00a3| = sina* and epicyclic behavior is assumed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002941_978-3-642-36279-8_13-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002941_978-3-642-36279-8_13-Figure1-1.png", + "caption": "Fig. 1 Figure 1(a) shows the parts used to construct truss structures. These parts include a node (bottom left), a member (right) and a module (top left). Members can be either horizontal beams or vertical columns. Figures 1(b) and 1(c) illustrate a quadrotor carrying a module and a column, respectively.", + "texts": [ + " These plans allow multiple robots to work concurrently on assembly while relying only on local information. This paper provides a description and theoretical evaluation of this algorithm, several heuristics, which assess its distributive nature, and analysis of the overall system performance for multiple quadrotors. In this section we briefly describe our previous work [8] upon which this paper is based. We also describe the capabilities of our experimental system. These truss structures are composed of two elementary components, nodes and members as shown in Fig. 1(a) similar in design to those presented in [3]. Each small cubic node can be attached to six members. Attaching a single node and member results in an additional component called a module also shown in Fig. 1(a). We will refer to members in their horizontal and vertical configuration as beams and columns, respectively. This design lends itself to the construction of cubic structures or extend to other structures like tetrahedral-octahedral structures by modifying the node design as shown later in Fig. 7(b). In our previous work, members and modules were placed in appropriately located pallets in known positions. Quadrotor robots with specially-designed grippers were able to pick up parts, transport them to the partially-built structure and assemble parts using simple trajectory controllers", + " In this paper, we overcome three limitations of the algorithm in [8] and extend the ideas to new cubic structures. First, we relax the assumption that planar stratum are required to be free of holes. Second, we allow for multiple quadrotors to assemble parts onto the structure concurrently. Finally, we only require local information for each quadrotor. In other words, quadrotors do not need to have global information about the state of the structure during assembly. We consider 2.5-D truss structures built from parts shown in Fig. 1(a). 2.5-D structures are those structures where next layer is fully contained by the previous layer. The 2.5-D restriction is purely a result of the part design. We also impose the constraint that parts cannot be inserted between previously placed parts like the deadlock illustrated in Fig. 2. Finally, we assume that each 2-D layer of the 2.5-D structure is connected. If there are multiple, disconnected components, we consider each component independently. We propose the Distributed Assembly Truss Structure (DATS) algorithm (Alg" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003889_isie.2013.6563767-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003889_isie.2013.6563767-Figure4-1.png", + "caption": "Fig. 4 Dynamic system", + "texts": [ + " It can measure very fast and with an extremely narrow field of view for the measurement when using laser light instead of sound. Fig. 3 is shown the flowchart of our system. Firstly constructing local map when the WMR receives laser data. In the normal condition, the WMR will be navigated to the goal by PSO algorithm, automatically. But in some conditions the WMR can\u2019t escape from a dead region. So we propose a wall following algorithm to find the way for WMR escapes out. III. NAVIGATION ALGORITHM AND DYNAMIC SYSTEMS To orientate the WMR, we define a coordinates system as shown in Fig. 4. The formula from 1 to 5 are show the definition of the xy coordinates and degree. The front of WMR is 0 degree, and dividing 360 degree into left and right 180 degree individually. Left direction is negative and right is positive, so they can be distinguished which direction should move on, clearly. The system inputs are the xy- coordinates of the WMR, degree between the WMR and the goal. The outputs are the turning angle and the step length that WMR moving. ( ) ( ) ( ( )) ( ) (1) ( ) ( ) ( ( )) ( ) (2) ( ) ( ) ( ) (3) where ( ) X- coordinates of WMR ( ) Y- coordinates of WMR ( ) Degree of WMR Distance between two wheels[m] Sampling time[s] Right wheel speed[m/s] Left wheel speed[m/s] ( ) Average speed of two wheels[m/s] ( ) Speed difference between two wheels[m/s] ( ) ( ) ( ) (4) ( ) ( ) ( ) (5) \u221a( ) ( ) (6) (3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002780_j.apm.2011.09.043-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002780_j.apm.2011.09.043-Figure1-1.png", + "caption": "Fig. 1. Schematic illustration of the gear-bearing system under nonlinear suspension.", + "texts": [ + " Nomenclature Cm damping coefficient of the gear mesh C1 damping coefficients of the supported structure for bearing 1 C2 damping coefficients of the supported structure for bearing 2 Cp damping coefficients of pinion Cg damping coefficients of gear C1p dimensionless parameter, C1p \u00bc m1 mp C01 dimensionless parameter, C01 \u00bc Kp1 K1 C2p dimensionless parameter, C2p \u00bc m2 mp C02 dimensionless parameter, C02 \u00bc Kp1 K2 Ep eccentricity ratio of pinion, ep/d Eg eccentricity ratio of gear, eg/d ep eccentricity of pinion eg eccentricity of gear Fx1 oil film force in the horizontal direction for bearing 1 Fy1 oil film force in the vertical direction for bearing 1 Fx2 oil film force in the horizontal direction for bearing 2 Fy2 oil film force in the vertical direction for bearing 2 fe1 and fu1 viscous damping forces in the radial and tangential directions for the center of journal 1 fe1 and fu2 viscous damping forces in the radial and tangential directions for the center of journal 2 f dimensionless parameter, f \u00bc mpg dKp1 fg dimensionless parameter, fg \u00bc Kp1g dmp fe, fu components of the fluid film force in radial and tangential directions Gpy, Ggy inertia forces in the vertical gear mesh direction for pinion and gear, Gpy \u00bc mpep \u20ach1 cos h1; Ggy \u00bc mgeg \u20ach2 cos h2 Km stiffness coefficient of the gear mesh K11, K12 stiffness coefficients of the springs supporting the two bearing housings for bearing 1 K21, K22 stiffness coefficients of the springs supporting the two bearing housings for bearing 2 Kp1 stiffness coefficients of pinion Kp2 stiffness coefficients of gear L bearing length Lpy, Lgy centrifugal forces in the vertical gear mesh direction for pinion and gear, Lpy \u00bc mpepx2 p sin h1;Lgy \u00bc mgegx2 g sin h2 m1 mass of the bearing housing for bearing 1 m2 mass of the bearing housing for bearing 2 mp mass of the pinion mg mass of the gear O1 geometric centers of the bearing 1 O2 geometric centers of the bearing 2 Oj1 geometric centers of the journal 1 Oj2 geometric centers of the journal 2 Og center of gravity of the gear Op center of gravity of the pinion p pressure distribution in the fluid film R inner radius of the bearing housing r radius of the journal s rotational speed ratio, s \u00bc X2 x2 n 1=2 s1 dimensionless parameter, s2 1 \u00bc C01C1ps2 s2 dimensionless parameter, s2 2 \u00bc C02C2ps2 Wcx dynamic gear mesh force in the horizontal direction, Wcx \u00bc Cm\u00f0 _Xp _Xg epX sin\u00f0Xt\u00de\u00de \u00fe Km\u00f0Xp Xg ep cos\u00f0Xt\u00de\u00de Wcy dynamic gear mesh force in the vertical direction, Wcy \u00bc Cm\u00f0 _Yp _Yg epX cos\u00f0Xt\u00de\u00de \u00fe Km\u00f0Yp Yg ep sin\u00f0Xt\u00de\u00de Xj, Yj horizontal and vertical coordinates, j \u00bc 1;2; j1; j2;p; g xj, yj Xj/c, Yj/c, j \u00bc 1;2; j1; j2;p; g a pressure angle a1 dimensionless parameter, a1 \u00bc K12d2Kp1 m1mp a2 dimensionless parameter, a2 \u00bc K22d2Kp1 m2mp b dimensionless parameter, b \u00bc Ep=16 bg dimensionless parameter, bg \u00bc Ep=16 n1 dimensionless parameter, n1 \u00bc C1 2 ffiffiffiffiffiffiffiffiffi K1m1 p n2 dimensionless parameter, n2 \u00bc Cp 2 ffiffiffiffiffiffiffiffiffiffi Kp1mp p n3 dimensionless parameter, n3 \u00bc Cm 2 ffiffiffiffiffiffiffiffiffiffi Kp1mp p n4 dimensionless parameter, n4 \u00bc Cg 2 ffiffiffiffiffi Kp1 mp q mg n5 dimensionless parameter, n5 \u00bc Cm ffiffiffiffiffi mp p 2mg ffiffiffiffiffi Kp1 p n6 dimensionless parameter, n6 \u00bc C2 2 ffiffiffiffiffiffiffiffiffi K2m2 p K dimensionless parameter, K \u00bc Km Kp1 Kg dimensionless parameter, Kg \u00bc Kmm2 p mg K2 p1 q mass eccentricity of the rotor / rotational angle, / \u00bc xt ui attitude angle of the line OiOji from the Xi-coordinate (see Fig. 1) X rotational speed of the shaft xn natural frequency, xn \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Kp1=mp p xg angular velocity of gear, xg = X/8 xp angular velocity of pinion, xp = X/4 h1 angular displacement of pinion, h1 = xpt h1 angular displacement of gear, h2 = xgt l oil dynamic viscosity e1 eccentricity ratio, e1 = e1/d e2 eccentricity ratio, e2 = e2/d bearings and used a suitable genetic algorithm to measure noise and model error. Theodossiades and Natsiavas [15] investigated dynamic responses and stability characteristics of rotordynamic systems interconnected with gear pairs and supported on oil journal bearings", + " Section 2 derives dynamic models for the gear-bearing system with a nonlinear suspension effect, turbulent flow assumption, strongly nonlinear gear mesh force and strongly nonlinear oil-film force. Section 3 describes the techniques used in this study to analyze the dynamic response of the gear-bearing system. Section 4 presents the numerical analysis results obtained for the behavior of the gear-bearing system under various operational conditions. Finally, Section 5 presents some brief conclusions. Fig. 1 presents a schematic illustration of the dynamic model considered in the present analysis. To simplify the whole dynamic system, the gyroscopic effect of rotor dynamics is neglected. Fig. 2 presents a schematic illustration of the dynamic model considered between gear and pinion. Og and Op are the center of gravity of the gear and pinion respectively; O1 and O2 are the geometric centers of the bearing 1 and bearing 2, respectively; Oj1 and Oj2 are the geometric centers of the journal 1 and journal 2, respectively; m1 is the mass of the bearing housing for bearing 1 and m2 is the mass of the bearing housing for bearing 2; mp is the mass of the pinion and mg is the mass of the gear; Kp1 and Kp2 are the stiffness coefficients of the shafts; K11, K12, K21 and K22 are the stiffness coefficients of the springs supporting the two bearing housings for bearing 1 and bearing 2; C1 and C2 are the damping coefficients of the supported structure for bearing 1 and bearing 2, respectively; Km is the stiffness coefficient of the gear mesh, Cm is the damping coefficient of the gear mesh, e is the static transmission error and varies as a function of time", + " (1)\u2013(4) can be performed as mp \u20acXp \u00fe Cp _Xp \u00fe Kp1 Xp Xj1 \u00bc Cm _Xp _Xg epX sin\u00f0Xt\u00de \u00fe Km Xp Xg ep cos\u00f0Xt\u00de ; \u00f05\u00de mp \u20acYp \u00fe Cp _Yp \u00fe Kp1 Yp Yj1 \u00bc mpepx2 p sin h1 Cm _Yp _Yg epX cos\u00f0Xt\u00de Km Yp Yg ep sin\u00f0Xt\u00de mpg; \u00f06\u00de mg \u20acXg \u00fe Cg _Xg \u00fe Kp2 Xg Xj2 \u00bc Cm _Xp _Xg egX sin\u00f0Xt\u00de Km Xp Xg eg cos\u00f0Xt\u00de ; \u00f07\u00de mg \u20acYg \u00fe Cg _Yg \u00fe Kp2\u00f0Yg Yj2\u00de \u00bc mgegx2 g sin h2 Cm _Yp _Yg egX cos\u00f0Xt\u00de \u00fe Km Yp Yg eg sin\u00f0Xt\u00de mgg; \u00f08\u00de The equations of motion of the center of bearing 1 (X1,Y1) and the center of bearing 2 (X2,Y2) under the assumption of nonlinear suspension (hard spring) are given by m1 \u20acX1 \u00fe C1 _X1 \u00fe K11X1 \u00fe K12X3 1 \u00bc Fx1; \u00f09\u00de m1 \u20acY1 \u00fe C1 \u20acY1 \u00fe K11Y1 \u00fe K12Y3 1 \u00bc m1g \u00fe Fy1; \u00f010\u00de m2 \u20acX2 \u00fe C2 _X2 \u00fe K21X2 \u00fe K22X3 2 \u00bc Fx2; \u00f011\u00de m2 \u20acY2 \u00fe C2 \u20acY2 \u00fe K21Y2 \u00fe K22Y3 2 \u00bc m2g \u00fe Fy2; \u00f012\u00de Considering the journal center Oj1 and Oj2, the resulting viscous damping forces in the radial and tangential directions are shown in Fig. 1. Applying the principles of force equilibrium, the forces acting at the center of journal 1, i.e. Oj1 (Xj1,Yj1), center of journal 2, i.e. Oj2 (Xj2,Yj2) [26] are Fx1 \u00bc fe1 cos u1 \u00fe fu1 sin u1 \u00bc Kp1\u00f0Xp Xj1\u00de=2; \u00f013\u00de Fy1 \u00bc fe1 sin u1 fu1 cos u1 \u00bc Kp1\u00f0Yp Yj1\u00de=2; \u00f014\u00de Fx2 \u00bc fe2 cos u2 \u00fe fu2 sin u2 \u00bc Kp2\u00f0Xg Xj2\u00de=2; \u00f015\u00de Fy2 \u00bc fe2 sin u2 fu2 cos u2 \u00bc Kp2\u00f0Yg Yj2\u00de=2; \u00f016\u00de where fe1 and fu1 are the viscous damping forces in the radial and tangential directions for the center of journal 1 respectively, and fe2 and fu2 are the viscous damping forces in the radial and tangential directions for the center of journal 2, respectively", + " (5)\u2013(16) can be expressed as e01 \u00bc b1dKp \u00f0yp y1 e1 sinu1\u00decosu1 \u00f0xp x1 e1 cosu1\u00desinu1 d1dKp \u00f0xp x1 e1 cosu1\u00decosu1\u00fe\u00f0yp y1 e1 sinu1\u00desinu1 4a3x c1d1\u00feb2 1 ; \u00f019\u00de u01 \u00bc 1 2 \u00fe dKp \u00f0yp y1 e1 sinu1\u00decosu1 \u00f0xp x1 e1 cosu1\u00desinu1 4xa3d1e1 ; b2 1dKp \u00f0yp y1 e1 sinu1\u00decosu1 \u00f0xp x1 e1 cosu1\u00desinu1 b1d1dKp \u00f0xp x1 e1 cosu1\u00decosu1\u00fe\u00f0yp y1 e1 sinu1\u00desinu1 4xa3e1d1 c1d1\u00feb2 1 ; \u00f020\u00de e02\u00bc b2dKp\u00bd\u00f0yg y2 e2 sinu2\u00decosu2 \u00f0xg x2 e2 cosu2\u00desinu2 d2dKp\u00bd\u00f0xg x2 e2 cosu2\u00decosu2\u00fe\u00f0yg y2 e2 sinu2\u00desinu2 4a4x\u00f0 c2d2\u00feb2 2\u00de ; \u00f021\u00de u02 \u00bc 1 2 \u00fe dKp\u00bd\u00f0yg y2 e2 sinu2\u00decosu2 \u00f0xg x2 e2 cosu2\u00desinu2 4xa4d2e2 b2 2dKp\u00bd\u00f0yg y2 e2 sinu2\u00decosu2 \u00f0xg x2 e2 cosu2\u00desinu2 b2d2dKp\u00bd\u00f0xg x2 e2 cosu2\u00decosu2\u00fe\u00f0yg y2 e2 sinu2\u00desinu2 4xa4e2d2 c2d2\u00feb2 2 ; \u00f022\u00de x00p \u00bc 2n2 s x0p 1 s2 xp x1 e1 cos u1 2n3 s x0p x0g Ep sin / K s2 \u00f0xp xg Ep cos /\u00de; \u00f023\u00de y00p \u00bc 2n2 s y0p 1 s2 yp y1 e1 sin u1 \u00fe b sin\u00f0/=4\u00de 2n3 s y0p y0g Ep cos / K s2 yp yg Ep sin / f s2 ; \u00f024\u00de x00g \u00bc 2n4 s x0g 1 s2 \u00f0xg x2 e2 cos u2\u00de \u00fe 2n5 s \u00f0x0p x0g Eg sin /\u00de Kg s2 \u00f0xp xg Eg cos /\u00de; \u00f025\u00de y00g \u00bc 2n4 s y0g 1 s2 \u00f0yg y2 e2 sinu2\u00de \u00fe bg sin\u00f0/=8\u00de \u00fe 2n5 s \u00f0y0p y0g Eg cos /\u00de \u00feKg s2 \u00f0yp yg Eg sin /\u00de fg s2 ; \u00f026\u00de x001 \u00bc 2n1 s1 x01 1 s2 1 x1 a3 s2 x3 1 \u00fe 1 2C1ps2 \u00f0xp x1 e1 cos u1\u00de; \u00f027\u00de y001 \u00bc 2n1 s1 y01 1 s2 1 y1 a3 s2 y3 1 \u00fe 1 2Coms2 \u00f0yp y1 e1 sinu1\u00de f s2 ; \u00f028\u00de x002 \u00bc 2n6 s2 x02 1 s2 2 x2 a4 s2 x3 2 \u00fe 1 2C2ps2 \u00f0xg x2 e2 cos u2\u00de; \u00f029\u00de y002 \u00bc 2n6 s2 y02 1 s2 2 y2 a4 s2 y3 2 \u00fe 1 2C2ps2 \u00f0yg y2 e2 sinu2\u00de f s2 ; \u00f030\u00de where a3 \u00bc lRL3 24c2Gz ; b1 \u00bc Z p 0 sin h cos h \u00f01\u00fe e1 cos h\u00de3 dh; c1 \u00bc Z p 0 cos2 h \u00f01\u00fe e1 cos h\u00de3 dh; d1 \u00bc Z p 0 sin2 h \u00f01\u00fe e1 cos h\u00de3 dh; a4 \u00bc lRL3 24c2Gz ; b2 \u00bc Z p 0 sin h cos h \u00f01\u00fe e2 cos h\u00de3 dh; c2 \u00bc Z p 0 cos2 h \u00f01\u00fe e2 cos h\u00de3 dh and d2 \u00bc Z p 0 sin2 h \u00f01\u00fe e2 cos h\u00de3 dh: Eqs. (19)\u2013(30) describe a non-linear dynamic system. In the current study, the approximate solutions of these coupled nonlinear differential equations are obtained using the fourth order Runge\u2013Kutta numerical scheme. In this study, the nonlinear dynamics of the gear-bearing system shown in Fig. 1 are analyzed using Poincar\u00e9 maps, bifurcation diagrams, the Lyapunov exponent and the fractal dimension. The basic principles of each analytical method are reviewed in the following sub-sections. The dynamic trajectories of the gear-bearing system provide a basic indication as to whether the system behavior is periodic or non-periodic. However, they are unable to identify the onset of chaotic motion. Accordingly, some other form of analytical method is required. In the current study, the dynamics of the gear-bearing system are analyzed using Poincar\u00e9 maps derived from the Poincar\u00e9 section of the gear system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.32-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.32-1.png", + "caption": "FIGURE 6.32 Full film hydrodynamic bearing motion from start-up.", + "texts": [ + " In particular, note the use of revolutions per second in the definition of speed in the Sommerfield number given in Equation 6.164. \ufffd \ufffd 227 Chapter | 6 Rotating Cylinders, Annuli, and Spheres a 2 \u03bcNsS \u00bc \u00f06:164\u00de c P where: S is the bearing characteristic number a is the journal radius, (m) c is the radial clearance (m) \u03bc is the absolute viscosity (Pa s) Ns is the journal speed (revolutions per second, rps) 228 Rotating Flow P = W/LD is the load per unit of projected bearing area (N/m2) W is the load on the bearing (N), D is the journal diameter (m), L is the journal bearing length (m). Consider the journal shown in Figure 6.32. As the journal rotates, it will pump lubricant in a clockwise direction. The lubricant is pumped into a wedgeshaped space, and the journal is forced over to the opposite side. The angular position where the lubricant film is at its minimum thickness ho is called the attitude angle, ho . 229 Chapter | 6 Rotating Cylinders, Annuli, and Spheres The center of the journal is displaced from the center of the bearing by a distance e called the eccentricity. The ratio of the eccentricity e to the radial clearance c is called the eccentricity ratio (\u03b5 = e/c), Equation 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001308_978-90-481-2505-0-Figure8.14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001308_978-90-481-2505-0-Figure8.14-1.png", + "caption": "Fig. 8.14 Scheme for the dynamic model derivation of the SCARA robot", + "texts": [ + " For transfer of the motion to the robot joints, belt transmission combined with gear trains is applied. The gear ratio is the same for both joints, i.e., N = 105/28. Each of the motors is equipped with an incremental encoder for measuring the shaft velocity and the direction of rotation, as well as its position. Other components of the system are two current controllers which in this case are analogue. For robot control, a special robotic version of DSP-2 control system, that can drive mechanisms with more than one degrees of freedom, was applied. Fig. 8.14. The dynamic model for the two degrees of freedom robot has the following known form: 1 11 12 1 1 2 21 22 2 2 T J J q C T J J q C = + && && (8.26) 8 Teaching of Robot Control with Remote Experiments 187 BookID 175907_ChapID 8_Proof# 1 - 14/4/2009 Here, iT is the sum of all torques acting on the i-th robot joint, iq&& is the acceleration of the i-th robot joint, iiJ is the inertia of the i-th joint, while ijJ and ijJ are coupling inertias. iC are torques caused by Coriollis and centrifugal forces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001298_s11431-010-3188-0-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001298_s11431-010-3188-0-Figure1-1.png", + "caption": "Figure 1 Relative position between toe and sensor array in locomotion (a), transformation between coordinates of sensor and coordinates of toe (b), relationship with reaction forces of toe, deflection angle and support angle (c).", + "texts": [ + " The locomotion morphology and 3-dimensional reaction forces of toes (tangential, radial and normal forces) were measured and collected by FLS when the geckos freely moved on walls and ceilings. The sensor array of FLS was made of 16 (2\u00d78) sensors and each sensor could measure 3-dimensional interaction toe forces. The positions of gecko\u2019s toes were recorded by a high speed camera at 215 frame/s (Mikrotron, MC1311, Germany) to obtain relative geometry positions between toes and load-carrier of sensor in locomotion [16, 17] (Figure 1(a)). The positions of toes were synchronized with reaction forces by lighting a LED in the view of the camera. Data regulation. When the geckos moved along the FLS, the relative position between the toes of the gecko and load-carrier of the sensor was always varying, thereby it was needed to transfer the toes reaction forces from the coordinate system of sensor (X1 Y1 Z1) to the coordinate system of toe (X2 Y2 Z2) (Figure 1(b)). The forces measured by sensor (FX FY FZ) could be converted into toe coordinate system ( T tF T rF T nF ) by \u03d5 \u03d5 \u03d5 \u03d5 \u23a1 \u23a4 \u2212 \u23a1 \u23a4\u239b \u239e \u23a2 \u23a5 \u239c \u239f \u23a2 \u23a5=\u23a2 \u23a5 \u239c \u239f \u23a2 \u23a5 \u23a2 \u23a5 \u239c \u239f \u23a2 \u23a5\u239d \u23a0 \u23a3 \u23a6\u23a2 \u23a5\u23a3 \u23a6 T t T r T n cos sin 0 sin cos 0 . 0 0 1 X Y Z F F F F FF (1) The shear force of toe T s ,F the reaction force of toe FT, the deflection angle of toe \u03b4 T and the support angle of toe \u03b2 T are calculated respectively by \u23a7 = + = +\u23aa \u23a8 = + + = + +\u23aa\u23a9 T T 2 T 2 2 2 s t T T 2 T 2 T 2 2 2 2 t n r r ( ) ( ) , ( ) ( ) ( ) , X Y X Y Z F F F F F F F F F F F F (2) \u03b4 \u03b4 \u03b2 \u03b2 \u23a7 = \u2208\u23aa \u23a8 = \u2208 \u2212\u23aa\u23a9 T T T T t r T T T T n s arctan( / ) (0 ,90 ), arctan( / ) ( 90 ,90 ), F F F F (3) where \u03c6 is the included angle between the coordinate system of sensor and the coordinate system of toe (Figure 1(c)). Data processing. The real contact area of all parts of a foot contacting with one sensor were selected from the high speed recording. The shear force of toe FT s , reaction force of toe FT, deflection angle \u03b4T and the support angle \u03b2T were calculated by eqs. (2) and (3). The maximum shear force of toe FT s and reaction force of toe FT were recorded. The corresponding reaction forces of toes (tangential force T tF , radial force T r ,F and normal force T nF ), the support angle \u03b2T and the deflection angle \u03b4T were selected respectively at the moment of the maximum reaction force of toe FT" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001614_978-3-642-03983-6_30-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001614_978-3-642-03983-6_30-Figure2-1.png", + "caption": "Fig. 2. The coordinate system with an earth frame {E} and a body frame {B}", + "texts": [ + " 4) The center of mass and body fixed frame origin coincide. These assumptions can be formed because of slower speed and lower altitude of the quad-rotor aircraft whose body is rigid having 6 DOF (Degree of Freedom) as compared to a regular aircraft. Under these assumptions, it is possible to describe the fuselage dynamics. Therefore, to mathematically illustrate the fuselage dynamics of the quad-rotor type aircraft, a coordinate system should be defined. The coordinate system can be divided into an earth frame {E} and a body frame {B} as shown in Fig. 2. The rotational transformation matrix between the earth frame and the body frame can be obtained based on Euler angles in Fig. 2 [1]. \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u2212 \u2212\u2212 +\u2212 == \u03c6\u03b8\u03c6\u03b8\u03b8 \u03c6\u03c8\u03c8\u03b8\u03c6\u03c8\u03c6\u03c8\u03c6\u03b8\u03c8\u03b8 \u03c8\u03c6\u03c8\u03b8\u03c6\u03c8\u03c6\u03c6\u03b8\u03c8\u03c8\u03b8 \u03c6\u03b8\u03c8 CCSCS SCSSCCCSSSSC SSCSCSCSSCCC RRRREB (1) where C and S indicate the trigonometric cosine and sine functions, respectively. The transformation of velocities between the earth frame and body frame can be derived from Eq. (1). \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 = \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 B B B EB vRv \u03c9 \u03c5 \u03c9 \u03c5 (2) Similarly, positions, forces, moments, accelerations and rotational velocities can be transformed based on REB between coordinate systems. In the body frame, the forces are presented as \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 = \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 = 4 3 2 1 1 1 1 1 0000 0000 F F F F F F F F zB yB xB B (3) Accordingly, in the earth frame, the forces can be defined as \u2211 =\u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u2212 + == \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 = 4 1 )( i iBEB zE yE xE E F CC SCSSC SSSCC FR F F F F \u03b8\u03c6 \u03c6\u03c8\u03c8\u03b8\u03c6 \u03c8\u03c6\u03b8\u03c8\u03c6 (4) Therefore, equations of motion in the earth frame are derived by the Newton\u2019s laws" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure14-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure14-1.png", + "caption": "Fig. 14. Drive unit of the prototype", + "texts": [ + " However, this structure is chosen for ease of fabrication. The important point is that the wires are fixed on the body at an angle; therefore, this structure is essentially equivalent to the one shown in the section II-C. If the four wires are pulled and released in turn, the prototype achieves helical rotating motion. We set the parameters as follows: Angle at which the wires are attached: \u03b1 = 84.9 [\u02da ] Distance between the wires and the body axis: r = 2.5 [mm] Assumed radius of the helix: a = 5.0 [mm] Assumed pitch of the helix: 2\u03c0b = 164 [mm] Figure 14 shows the drive unit of the prototype. The actuators are set outside of its body. Although the prototype is not self-contained, this actuator arrangement is practical for an endoscope for humans because it yields a simple body. Four servomotors are mounted on the base of the drive unit. A pulley with a diameter of 50 mm is fixed on the shaft of each motor. The wires of the prototype are wound on the pulleys, which control the amount by which the wires are pulled. If the end of the helical body is fixed on the drive unit rigidly, the body swings around, and its posture is unsteady" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000945_978-1-4020-8600-7_5-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000945_978-1-4020-8600-7_5-Figure4-1.png", + "caption": "Fig. 4 Singular configurations of the 5-dof parallel manipulator.", + "texts": [ + " The singular configurations of the parallel manipulator can be algebraically derived from Eq. (27). Expanding Eq. (27) leads to an algebraic equation in terms of the five joint variables of the RRPRR leg (all joint screws of the parallel manipulator are expressed as functions of these five variables). The solutions of this equation give the singular configuration of the manipulator. The expanded equation is not listed here because of the limited space. Several singular configurations have been identified. Two types of singular configurations are shown in Figure 4. The uncontrollable motion of the moving platform for the first singular configuration (Figure 4a) is a pure rotation, which axis intersects all four legs, is parallel to Geometric Algebra Approach to Singularity of Parallel Manipulators two R joint axes of the RRPRR leg and perpendicular to P joint axis of the RRPRR leg. In the second singular configuration (Figure 4b) the uncontrollable motion is a general screw motion. The presented approach proves to be effective in determining the singularity condition for parallel manipulators with limited mobility. This approach is applied to two parallel manipulators, which singular configurations are obtained. It has been shown that the equation for the singularity (the condition for singularity) involves the screws which represent all and only passive joints of the manipulators. This geometric algebra approach provides a good geometrical insight and efficiency in dealing with robot kinematics and singularity of parallel manipulators with fewer than six degrees of freedom" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000348_taes.2008.4516988-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000348_taes.2008.4516988-Figure1-1.png", + "caption": "Fig. 1. Longitudinal forces and moment acting on missile.", + "texts": [ + ", post burnout, no roll rate, zero roll angle, no sideslip, and no yaw rate. The nonlinear equations of motion for a rigid airframe reduce to two force equations, one moment equation, and one kinematic equation _U+ qW = X FBX m (1) _W\u00a1 qU = X FBZ m (2) _q= X MY IY (3) _\u03bc = q (4) where U and W are components of velocity vector ~VT along the body-fixed X and Z axes; \u03bc is the pitch angle; q is the pitch rate about the body Y axis; m is the missile mass. The forces along the body-fixed coordinates and moments about the center of gravity are shown in Fig. 1. The forces and moment about the center of gravity are described byX FBX = Lsin\u00ae\u00a1Dcos\u00ae\u00a1mg sin\u03bc (5)X FBZ =\u00a1Lcos\u00ae\u00a1D sin\u00ae+mg cos\u03bc (6)X MY = M\u0304 (7) where \u00ae is the angle of attack, L denotes aerodynamic lift, D denotes drag, and M\u0304 is the total pitching moment. The lift, drag, and pitching moment are as follows L= 1 2\u00bdV 2 T SCL, D = 1 2\u00bdV 2 T SCD, M\u0304 = 1 2\u00bdV 2 T SdCm (8) where \u00bd is the air density and VT is the missile speed, i.e., VT = p U2 +W2. Note that the normal force coefficient is used to calculate the lift and drag coefficients: CL =\u00a1CZ cos\u00ae, CD = CD0 \u00a1CZ sin\u00ae (9) where CD0 is the drag coefficient at zero angle of attack" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001797_aqtr.2010.5520862-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001797_aqtr.2010.5520862-Figure1-1.png", + "caption": "Figure 1. Kinematics model of the robot", + "texts": [ + " MOBILE ROBOT SYSTEM The used robot is a full autonomous wheeled robot intended for indoor environment. It has two identical parallel, wheels (attached to both sides of the vehicle) which are controlled by two independent DC gear motors. Also it assumed that each wheel is perpendicular to the ground and the contact between the wheels and the ground is pure rolling and non-slipping. The velocity of the center of mass of the robot is orthogonal to the wheels\u2019 axis - L. The center of mass of the mobile robot (x, y) is located in the middle of the axis connecting the wheels (L), likes in Fig. 1. This robot type is relatively easy to model and build. The angular velocities ( l and r) of the two wheels are independently controlled. This mobile robot is an under actuated system, because it has two inputs (translational velocity \u2013 vc and angular velocity \u2013 \u03c9c) and three outputs (center positions x, y and heading angle \u03b8 of the mobile robot on two dimensional Cartesian workspace) [16]. The equations that are used to build a MatlabSimulink model of the robot are given by relation (1). \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 \u22c5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 \u2212 = \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 l r x x v v L 1 L (t) (t)v (t)v 1 sin 2 1sin 2 1 cos 2 1cos 2 1 \u03b8\u03b8 \u03b8\u03b8 \u03b8 (1) This model is the simplest and the most suitable for a smallsized and light, battery-driven autonomous robot" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001639_biorob.2010.5628009-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001639_biorob.2010.5628009-Figure6-1.png", + "caption": "Fig. 6. Change in shape of a cylindrical body when a wire aligned with the body axis is contracted.", + "texts": [ + " Next, when wire C is contracted while B is elongated, the body rotates again. When this motion is repeated, the body makes a rotating motion. Therefore, rotating motion can be achieved by contraction and elongation of the four wires. This body can also roll across the ground. The wires are aligned with the body axis in the above mentioned case. Next, we consider the case where the wire is attached at an angle to the body axis. If the wire is aligned and the wire is contracted, the shape of the body becomes a circular arc, as shown in Fig. 6. However, if the wire is fixed at an angle to the axis of the body and the wire is contracted, the shape of the body becomes helical, as shown in Fig. 7. Therefore, if four wires are attached to the body at an angle and the wires are contracted and elongated in turn, the body as a whole rotates on its axis. This is helical rotating motion. When this motion is achieved in a pipe, every part of the body rolls on the inner wall of the pipe, as shown in Fig. 8. Thus, helical rotating motion generates a propulsive force in a pipe" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000076_j.robot.2008.01.002-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000076_j.robot.2008.01.002-Figure4-1.png", + "caption": "Fig. 4. Foot point coordinate of a leg.", + "texts": [ + " 3(b), and the failed leg moves by swinging joint one as shown in Fig. 3(c), making the track of foothold positions in the shape of an arc. When the leg trajectory is a straight-line, the above characteristics of locked joint failures makes the failed leg have only one possible foothold position. Hence, in our study, a given configuration of the locked joint failure can be represented by a point on the workspace. We define such a point as the coordinate on the axis that has the direction of the leg trajectory and the origin at the center point Ci . Fig. 4 shows our definition, where ri is the position of a failed leg on the designated axis. Note that ri > 0 at point A, i.e., if the failed leg is placed in front of Ci , while ri < 0 at point B. If \u03b1 = 0, the length of leg trajectory segment in the workspace is Rx . In general, if \u03b1 \u2265 0, the segment, denoted by R\u03b1 , is derived as R\u03b1 = { Rx/ cos \u03b1 0 \u2264 \u03b1d < \u03b11 Ry/ sin \u03b1 \u03b11 \u2264 \u03b1d < 90\u25e6. (1) Under this definition, R0 = Rx and \u2212R\u03b1/2 < ri < R\u03b1/2. We will use this definition to represent supporting positions of normal legs as well", + " For simplicity of analysis, we begin with the assumption that a locked joint failure occurs to leg 1 throughout this paper. Fault-tolerant gaits for failures at other legs can be derived by symmetry of the quadruped robot. Fig. 5 shows the previously developed fault-tolerant periodic gait [18] for a quadruped robot having straight-line going, i.e., with \u03b1 = 0, over even terrain. The gait in Fig. 5 is a kind of the discontinuous gait characterized by the sequential motion of legs and the robot body. r1 in the figure denotes leg 1\u2019s foot point defined in Fig. 4. Note that r1 is in the range of 0 < r1 < Rx/2 in Fig. 5, or leg 1 is on the right half of the working area in the forward direction. The leg sequence for the case of \u2212Rx/2 < r1 \u2264 0 is the same as that of the gait shown in Fig. 5 with a slight modification of the initial foot points of the three normal legs (refer to [18]). The key feature of this gait is that the failed leg is kept its foot point constant with respect to the center of gravity throughout walking for avoiding any malfunction caused by the locked joint" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002232_j.mechmachtheory.2012.02.005-Figure4-1.png", + "caption": "Fig. 4. Overlapping the gears with translated wheel: a \u2013 overlapping scheme, b \u2013 overlapped gear.", + "texts": [ + " Then, using the \u201cGoal Seek\u201d tool from Excel, the final value S1f which satisfies the function f(S1, \u03c62)=0 is searched; in this way, the correlation S1/\u03c62 is obtained, for each value of the angle \u03c62. Based on the values S1/\u03c62 determined in this way, the rolling radius of the sector r2r (6) and the transmission ratio it (1) are obtained. The gearing process is analyzed on the base of fictive gear with centers O1\u2212O2, with the pressure angle \u03b1 and gearing pole C, which depend on S1/\u03c62. This fictive gear can overlap the initial reference position, O2\u2192O2 0 and O1\u2192O1 r , obtaining in this way the gear of centersO1 r\u2212O2 0 (Fig. 4a)with the distance \u201ca\u201d between centers. Pushing thewheel 1with\u0394a=a\u2212a0 on the directionO2 0\u2212O1 0 will determine the clearance between teeth (Fig. 4b), the clearance jn on the normal axis n\u2013n of the fictive gear of pressure angle \u03b1 being the distance A\u2032E\u2032 on the line of action, given by relation: jn \u00bc AE\u2212 AA0 \u00fe EE0 \u00bc a sin\u03b1\u2212 rb1 \u00fe rb2 \u03b1 \u00fe \u03b80\u00f0 \u00de; \u03b80 \u00bc tg\u03b10\u2212\u03b10: \u00f09\u00de In this equation, it is known that the length EE0 or AA0 from the line of action is equal with the arc from the base circle, and \u03b80= inv(\u03b10) is the involute angle. Note: relations (3)\u2013(9) can be generalized also for the solution shown in Fig. 1c, considering negative values for eccentricity \u201ce\u201d" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001021_icelmach.2010.5607938-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001021_icelmach.2010.5607938-Figure1-1.png", + "caption": "Fig. 1. Experimental test stand.", + "texts": [ + " In order to validate the proposed approach for diagnosis of eccentricities in IFIMs, a series of laboratory tests were carried out on a commercial cage motor. The inverter used was a Micromaster 420 (Siemens). The characteristics of both, the motor and the inverter, are given in the appendix. Two different conditions were considered: healthy machine and machine with eccentricity. The eccentricity was achieved by means of an irregular coupling of the motor with a brake and by additionally placing a L-shaped steel plate under one of the supports of the platform. Fig.1 shows the test rig. The effects of this artificially created eccentricity on the IFIM stator steady state current have been deeply analyzed in [11]. As mentioned in the introduction, the presented methodology can be applied in case of failure of the FFT method (mainly when diagnosing an IFIM which rarely works in steady state conditions, being this regime too short in case of appearance). The diagnosis methodology first step consists of the capture of a transient current of the electric machine being diagnosed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001895_978-3-642-32448-2_11-Figure11.4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001895_978-3-642-32448-2_11-Figure11.4-1.png", + "caption": "Fig. 11.4 Dexel representation of a workpiece in three directions in space", + "texts": [ + " Based on the chip geometry the milling forces are calculated using a standard model presented in [3]. While other methods for cutter workpiece engagement simulations are introduced in [12, 13, 14, 15] the dexel representation used in this work enables an efficient computation of the engagement condition. Especially in applications of milling with industrial robots the accuracy of the dexel-based method is considered to be appropriate. In order to increase the calculation accuracy, [16] recommends the usage of a multi-dexel model for the representation of the workpiece (cf. Fig. 11.4). Thereby, the description of a single dexel is made by line equations, t= +p d s Here, p is a point on the line, d is the line direction, t is the scaling factor and s the starting point in space. The workpiece is modeled using a multi-dexel representation compromising dexel directions in x, y and z. The discretization of the workpiece is defined by the dexel distances dx, dy and dz. In order to receive a sufficient accuracy of the force calculation the relation kd between the dexel discretization and the tool radius R is considered, 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003759_msec2015-9396-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003759_msec2015-9396-Figure2-1.png", + "caption": "Figure 2. Illustration of the filament extruder", + "texts": [ + " As shown in Figure 1, thermoplastic filament is fed into the extrusion nozzle by the two filament drive wheels then extruded through the extrusion nozzle that trace the object model\u201fs cross sectional geometry. The resistance heater keeps the thermoplastic filament at a temperature just above its melting point. The thermoplastics harden immediately after flowing through the extrusion nozzle and bond to the layer below. After one layer is complete, the build platform lowers slightly to make way for the next layer. In the same principle, each single layer is built on atop the other until the object model is complete. A preliminary extrusion process as illustrated in Figure 2 is required to produce thermoplastic filaments for use in FDM 3D printers. Thermoplastic pellets as shown in Figure 3 are fed into the pellet hopper. The pellets flow by gravity from the pellet hopper down into the barrel and fill the annular space between the screw and the barrel. Since the barrel is stationary and the screw is rotating, the frictional force acts on the pellets, the barrel, and the screw surface. As the pellets moving downwards, they are heated up and melted by the resistance heater and the frictional heat generated, and then forced out of the nozzle by the screw to form a continuous filament" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003801_1.4024294-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003801_1.4024294-Figure1-1.png", + "caption": "Fig. 1 A humanoid redundant manipulator shown with its joint angles", + "texts": [ + "24) The most typical redundant manipulator is the human arm, which is modeled to have seven degrees of freedom for all practical purposes as mentioned in several sources including [26]. Therefore, the demonstration manipulator considered in this paper is selected to be a humanoid manipulator. A humanoid manipulator has spherical joint arrangements at its shoulder and wrist points (S and R) and a revolute joint at its elbow point (E). A spherical joint arrangement consists of three successive revolute joints with concurrent but noncoincident axes. A general view of a humanoid manipulator is shown in Fig. 1. The angles of the second, fourth, and sixth joints are indicated directly on the sketch of the manipulator. The angles of the other joints together with the relevant unit vectors are indicated on the auxiliary sketches below the main sketch. In this paper, the torso is assumed to remain stationary in an upright position. The reference frame attached to the torso is F 0\u00f0O\u00de; where O is its origin on the axis of the neck. The reference frames attached to the moving links are fitted in such a way that they are all parallel to F 0\u00f0O\u00de when the joint angles are zero", + " 5, AUGUST 2013 Transactions of the ASME Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/27/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 3.1 Forward Kinematics in the Position Level. Let C\u00f0i;j\u00de be the matrix that represents the orientation of the jth link with respect to the ith link. It also functions as the component transformation matrix from the link frame F j to the link frame F i. Considering the axes and the angles of the joints indicated in Fig. 1, the orientation matrices between the successive links can be expressed as follows: C\u00f00;1\u00de \u00bc e~u1h1 ; C\u00f01;2\u00de \u00bc e~u2h2 ; C\u00f02;3\u00de \u00bc e~u1h3 C\u00f03;4\u00de \u00bc e~u2h4 C\u00f04;5\u00de \u00bc e~u1h5 ; C\u00f05;6\u00de \u00bc e~u2h6 ; C\u00f06;7\u00de \u00bc e~u1h7 9= ; (3.1) In Eq. (3.1), {h1, h2, h3} is the set of the shoulder angles, h4 is the elbow angle, and {h5, h6, h7} is the set of the wrist angles. The orientation matrices are expressed exponentially with a general notation as e~n/. Here, e~n/ is briefly called an exponential rotation matrix as in Ref" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003547_1.4005013-Figure15-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003547_1.4005013-Figure15-1.png", + "caption": "Fig. 15 Experimental setup with multichannel and partitions", + "texts": [ + " Since the kinetic energy of balls is dissipated by the collision, the self-excited oscillation ceased and the speed range where self-excited oscillations occur becomes narrower. If the restitution coefficient is made smaller by attaching; for example, rubber sheets to the partitioned, the self-excited oscillation can be suppressed easier. From the discussion above, it is clear that the recommended ball balancer is a multichannel ball balancer with partitions. In this chapter, the effect of a multichannel ball balancer is studied in experiments. 7.1 Experimental Setup. Figure 15 shows the experimental setup of a multichannel ball balancer with partitions. There are four partitions in each channel and one ball is put in each chamber. Angular positions of the partitions on the upper and lower sides differ by 45 as shown by the full and the broken lines. (Although the reason was not clear, this obtained better results than the case where the upper and lower partitions were inserted at the same angular positions.) The closed and shaded Journal of Vibration and Acoustics APRIL 2012, Vol" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003288_1.4769758-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003288_1.4769758-Figure7-1.png", + "caption": "FIG. 7. Capillary attraction on an arbitrary curved interface. (\u03be , \u03b7) are the coordinates used to describe the background curved interface, and (x, y) are the local coordinates at the location of the particles.", + "texts": [ + " The gravitational curvature contribution, F\u2217 c is clearly small compared to the weight of the particle, F\u2217 w. It is also interesting to note that the shape of the curvature contribution, F\u2217 c obtained here has a striking similarity with the variation of gradient of curvature for this interface. We now treat the case of attraction between horizontal cylinders floating on an interface with a finite background curvature, the motivation for which was given in the Introduction. Consider two cylinders on an interface with a known background curvature as shown schematically in Figure 7, where the dashed curve represents the undisturbed interface in the absence of particles. We use the same perturbed Young-Laplace equation, (40), as used in Sec. V but with the origin of the local coordinate system (x, y) now lying on the undisturbed interface midway between the two particles. The coefficients a(\u03be ) and b(\u03be ) in Eq. (40) are evaluated at the origin of the local coordinate system, This article is copyrighted as indicated in the article. Reuse of AIP content is subject to the terms at: http://scitation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002612_j.csefa.2013.05.002-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002612_j.csefa.2013.05.002-Figure6-1.png", + "caption": "Fig. 6. As illustrated in the schematic diagram (Fig. 5), damage at roller tip and deep groove on the cone were observed on the premature failed bearing.", + "texts": [ + " This is possible due to decomposition of retained austenite in service. The retained austenite transforming to work induced martensite causes compressive residual stress at the crack tip that is helpful in delaying the propagation of crack. This is considered beneficial for rolling contact fatigue. The transformation of RA to martensite is responsible for the higher hardness of 64\u201365 HRc at and near surface regions than the core. The above proposed mechanism is in concurrence with the failure observed in the samples as shown in Fig. 6. This misalignment may have been caused at the time of installation or during operation where the bearing would have experienced any excessive load. The failure is not on account of any defect in material or heat treatment. This failure is predominantly due to the misalignment, which caused uneven load distribution on the roller. While the intentionally designed crowning of the rollers may accommodate very small misalignment, it cannot prevent damage when the misalignment is excessive. In the present case, the misalignment may have been excessive" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003835_j.arabjc.2013.08.012-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003835_j.arabjc.2013.08.012-Figure4-1.png", + "caption": "Figure 4 3D response surface plot showing the interaction of the lipase concentration and molar ratio on butyl butyrate formation.", + "texts": [ + " A ratio greater than 4 is desirable. Adequate precision measures the signal to noise ratio. A relatively lower value of the coefficient of variation (CV = 12.53) indicated a good precision and reliability of the experiment. To determine the optimal levels of each variable, the graphical representation of the regression equation in the form of three dimensional response surface plots was constructed by plotting the response on the Z-axis against the two independent variables (molar ratio and lipase concentration) as shown in Fig. 4. An increase in esterification yield was observed when the lipase concentration or molar ratio increased, indicating the interaction of the two in enhancing the yield up to a certain limit. The esterification yield (butyl butyrate formation) obtained in this study was found to be promising based on the literature reports of different lipase-catalyzed synthesis of esters using free and immobilized enzyme systems. Babu and Divakar (2001) studied the synthesis of esters of anthranilic acid using Please cite this article in press as: Salihu, A" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001355_14644193jmbd187-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001355_14644193jmbd187-Figure1-1.png", + "caption": "Fig. 1 Schematic diagram of rolling element bearings", + "texts": [ + " The rolling elements are positioned symmetrically such that their moving parts are in synchronization. To determine how the nonlinear bearing forces act on the system, the implicit type numerical integration technique Newmark-\u03b2 [30] with the Newton\u2013Raphson method is used to solve the non-linear differential equations iteratively. The results obtained from a large number of numerical integrations are mainly presented in the form of fast Fourier transformation (FFT) and Poincar\u00e9 maps. A schematic diagram of a rolling element bearing is shown in Fig. 1. For investigating the structural vibration characteristics of a rolling element bearing, a model of the bearing assembly can be considered as a non-linear spring mass damper system. Elastic deformation between races and rollers gives a nonlinear force\u2013deformation relation, which is obtained by Hertzian theory. In the mathematical modelling, the rolling element bearing is considered as a spring mass system and rolling elements act as a non-linear contact spring, as shown in Fig. 2. Since the Hertzian forces arise only when there is contact deformation, the springs are required to act only in compression" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003054_rob.21462-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003054_rob.21462-Figure9-1.png", + "caption": "Figure 9. Terrain inclination angles at Rij . The pseudoplane is calculated from the wheel contact points.", + "texts": [ + " Each index can be computed based on the C2DEM representation. A rover experiences relatively high wheel slippage when it climbs up or traverses a sloped terrain. This slippage is due to the traction load from the gravity, which becomes significant on the sloped terrain. Additionally, rollover of a rover traversing a steep slope is a mobility hazard. The terrain inclination index is employed for the cost function to represent such risks on a sloped terrain. The terrain inclination index is divided into two axes, namely roll and pitch of the rover (Figure 9). Each axis is geometrically calculated as an angle between the inertial coordinate and a pseudoplane composed of the wheel contact points at a projection region of the rover Rij : Multiple Journal of Field Robotics DOI 10.1002/rob terrain inclinations can be calculated from each subset of three contact points between the multiple wheels. The terrain inclination index is the largest inclination between these values: Roll : x ij = max[\u03b8x (Rij )], (3) Pitch : yij = max[\u03b8y(Rij )]. As illustrated in Figure 10, the projection region Rij is determined with the dimension of the rover" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003569_melcon.2012.6196611-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003569_melcon.2012.6196611-Figure1-1.png", + "caption": "Fig. 1. Twin Rotor Multi-In", + "texts": [ + " To FLC and the wellbeen proposed in ility of the FSMC his technique has plications [1], [9], s that it has fewer using the SMC more robustness sturbances. the developing of arameters for the is problem is the ms. Optimization eir capabilities of trained, or even )-hard problems. s one can found in A), Ant Colony timization (PSO), d to tune the fuzzy rs and the sliding r angle positions nized as follows: ction 2. The PSO- in Section 3. The in Sections 4 and The TRMS is a laboratory s Instruments Limited for control Fig. 1, this system is charact nonlinear dynamics, where som measurements, and hence this challenging engineering problem beam pivoted on its base in s freely in both horizontal and ve of the beam, there are two rot driven by direct current (DC) m with a weight is fixed at the piv 1). The control objective is to set desired positions (the pitch an quickly and accurately. All t parameters are given in the follo \u2212 Fh/v( h/v): Two nonlinear dependence of propeller th speeds h/v, \u2212 Ph/v(uh/v): Two nonlinear inpu dependence of tail/main DC tail/main input voltage uhh/vv, \u2212 lt: Length of the tail part of b \u2212 lm: Length of main part of bea \u2212 h/v: Horizontal/vertical positi \u2212 Jv: Moment of inertia about h \u2212 Jh( v0): Moment of inertia abo \u2212 kh/v: Positive constants, \u2212 Ttr,Tmr : Time constants of tail \u2212 ktr/mr: Tail/main motor propell \u2212 Jtr/mr: Moment of inertia in subsystem (kg\u00b7m2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003725_icuas.2015.7152366-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003725_icuas.2015.7152366-Figure5-1.png", + "caption": "Fig. 5. Computation of induced relative air velocity v\u221e,i and induced angle of attack \u03b1i by means of superposition of flows assumption.", + "texts": [ + " An appropriate model is therefore necessary and described in the following. We follow the guidelines settled in [9] and define two different wing sections as illustrated by Fig. 4. We assume one of them to be unaffected by the propulsion wake and is modeled according to equations 20, 21 and 22. The other will be affected by a slip-stream wake profile and its effect on the aerodynamic equations is modeled by means of a induced free-stream velocity v\u221e,i and an accordingly induced angle of attack \u03b1i as illustrated by figure 5. The geometry of the model suggests superposition over the air flows and the following modification in the angle of attack equation (see equation 14)( cos\u03b1i sin\u03b1i ) = \u22121 v\u221e,i ([ cos\u03b8 \u2212sin\u03b8 sin\u03b8 cos\u03b8 ]( wn \u2212 vn wd \u2212 vd ) \u2212 ( vi 0 )) (27) where vi is the induced velocity of slip-stream at the downstream side of propeller disk which, by means of the propeller momentum theory for forward helicopter flight [10], is a solution of v2i + v\u221ecos\u03b1vi \u2212 T 2\u03c1Sp = 0 (28) where Sp denotes propeller disk area. Therefore, we write vi = 1 2 [\u221a (v\u221ecos\u03b1)2 + 2T \u03c1Sp \u2212 v\u221ecos\u03b1 ] (29) where the \u03b1-based terms can be again passed-by by substitution of equation 14" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002975_physreve.87.053021-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002975_physreve.87.053021-Figure1-1.png", + "caption": "FIG. 1. (Color online) (a) Experimental apparatus. (b) Sampled images of falling parallelograms. The two views show that the card tumbles and falls in the shape of a helix. (c) A parallelogram and its length L, width W , internal angle \u03b3 , center of mass (com), diagonal axis x, chord axis y, and tumbling axis (schematic).", + "texts": [ + " In what follows, we report detailed measurements of the kinematics and aerodynamic forces of the falling parallelograms by varying parameters, such as mass density, aspect ratio, and internal angle. The experimental aerodynamic forces and torques may guide future development of 3D aerodynamic models of falling objects. II. EXPERIMENTAL APPARATUS AND MEASUREMENT PROCEDURE The experiments are performed in still air. A schematic diagram of the experimental setup and arrangement of the apparatus are shown in Fig. 1. In this study, flat, parallelogram shaped, white cards of length L, width W , thickness d (with d [W,L]), and internal angle \u03b3 are dropped in still air with their long axis horizontal and acute axis vertical from an approximate height of 1.5 m. The long axis is parallel to the card length and passes through the center of mass, and the acute axis is perpendicular to the long axis and also passes through 053021-11539-3755/2013/87(5)/053021(7) \u00a92013 American Physical Society the center of mass. The card parameters, such as aspect ratio L/W , internal angle \u03b3 , and mass density \u03c1s , can influence the falling trajectory", + " Other parameters which influence the dynamics of the cards in question are the mass density of the card \u03c1s = 1.2 \u00d7 103 kg/m3, the density of air \u03c1a = 1.23 kg/m3, the gravitational acceleration g = 9.8 m/s2, and the dynamic viscosity of air \u03bc = 1.73 \u00d7 105 (N-s/m2). Four dimensionless parameters relevant in this analysis are the Reynolds number Re = \u03c1a w2/\u03bc, the aspect ratio L/W, the buoyancy number \u03c1s/\u03c1a \u2212 1, and the thickness-to-width ratio of the card, d/W . The last two can be combined to obtain the dimensionless moment of inertia I \u2217 = \u221a \u03c1sd \u03c1aW . In Fig. 1, the experimental apparatus includes a high-speed camera (Vision Research, Phantom v5.0), a mirror of size 76 \u00d7 51 cm which is mounted at 45\u25e6 from the horizontal, and two black backgrounds of which one is placed on the opposite side of the mirror and the other is placed horizontally at 2 m above the mirror. Both backgrounds are illuminated homogeneously to obtain good contrast. The sampling rate needs to be sufficient to enable our subsequent analysis of the card kinematics. Our results indicate that 1000 frames/s were sufficient to resolve the motion of the card, and therefore the trajectories of the falling cards are recorded at this frame rate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003294_jjap.51.06fl17-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003294_jjap.51.06fl17-Figure1-1.png", + "caption": "Fig. 1. (Color online) Fabrication process of metallic movable microparts by the combination of two-photon microfabrication and electroless plating. The movable microparts supported with anchors are fabricated by twophoton microfabrication. After electroless plating, the anchors are removed by laser ablation.", + "texts": [ + " In our method, since the movable microparts are separated from the substrate by the anchors, even the bottom area of the movable microparts can be metalized. The anchor supports are removed by laser ablation with a femtosecond pulsed laser beam. The combination of the anchor supporting method and laser ablation enables us to make 3D metallic movable microparts. To metalize movable microparts produced by two-photon microfabrication, we propose a fabrication process using electroless plating and laser ablation. Figure 1 shows the fabrication process of a metallic movable micropart. The detail of the process is as follows: (Step 1) A microrotor supported with anchors is fabricated on a glass substrate by two-photon microfabrication. (Step 2) The movable micropart is metalized by electroless copper plating. Since the movable micropart is suspended by the anchors, electroless plating is performed on the entire movable micropart. (Step 3) Finally, the anchors are removed from the movable micropart by laser ablation with a femtosecond pulsed laser beam" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000945_978-1-4020-8600-7_5-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000945_978-1-4020-8600-7_5-Figure1-1.png", + "caption": "Fig. 1 Four-bar mechanism (one-dof parallel manipulator).", + "texts": [ + " Using the language of the geometric algebra, the condition of singularity for the parallel manipulator with less than six degrees of freedom (but with dummy joints) can be expressed in the following way Da1 \u2227 \u00b7 \u00b7 \u00b7 \u2227Dak \u2227Dd1 \u2227 \u00b7 \u00b7 \u00b7 \u2227Ddq = 0, k + q = 6, (15) where Dai is a dual vector (grade 1-blade) associated to the ith active joint andDdi is a dual vector (grade 1-blade) associated to the ith dummy joint. In this case each leg has a full mobility. Here the dummy joints are considered as active but locked. In the following sections, the approach is applied to two particular parallel manipulators. Firstly, in order to illustrate the approach, it is applied to a very simple example, i.e. to the four-bar mechanism (Figure 1a), whose singular configurations are known (e.g., Zlatanov et al., 2002). It is considered as a planar parallel manipulator with two RR-legs and the coupler as a moving platform. The mechanism has one driven joint with the joint axis 1S1. In order to have full mobility, we suppose that one dummy joint is added to each leg. Then, the duals corresponding to the active and dummy joints for the first and the second leg, respectively, are as follows: 1D1 = (1S2 \u2227 1Sd \u2227 e126)I \u22121 6 ; 1Dd = (1S1 \u2227 1S2 \u2227 e126)I \u22121 6 , (16) 2Dd = (2S1 \u2227 2S2 \u2227 e126)I \u22121 6 , (17) where e126 = e1 \u2227 e2 \u2227 e6 is a 3-blade representing the restricting subspace, i", + " The blade from Eq. (18) is a blade of non-freedom for the first leg. In this case, the wrenches of constraint associated with the dummy joints for the first and the second leg, respectively, are 1Cd = 1D\u0303d , 2Cd = 2D\u0303d (derived from Eqs. (16) and (17)) and the third one can be obtained by factoring the 2-blade from Eq. (18). In this case the wrenches of constraints are pure forces, Notice, that the two constraint forces 1Cd and 2Cd , associated with the dummy joints, are unique (along the legs, Figure 1a). The condition for singularity of the manipulator can be written as 1D1 \u2227 1Dd \u2227 2Dd = 0 or 1D\u03031 \u2227 1D\u0303d \u2227 2D\u0303d = 0. (19) Again, applying the identities of the geometric algebra and keeping in mind Eq. (18), the left-hand blade (the singularity condition) from Eq. (19) becomes [c(1S2 \u2227 e126)I \u22121 6 ] \u2227 [(2S1 \u2227 2S2 \u2227 e126)I \u22121 6 ] = \u2212c(2S1 \u2227 2S2 \u2227 e126 \u2227 1S2)I \u22121 6 e126I \u22121 6 = c(2S1 \u2227 2S2 \u2227 1S2 \u2227 e126)e126 = 0. (20) Therefore, bearing in mind that e1 \u2227 e2 \u2227 e6 = 0, it is clear from Eq. (20) that the condition for singularity can be written as 2S1 \u2227 2S2 \u2227 1S2 = 0. (21) Eq. (21) implies that the mechanism is in singular configuration if the three lines (joint axes, which are parallel) are linearly dependent, i.e., lie in a single plane, defined by any two of the lines (Figure 1b). Eq. (21) involves only the screw axis of the passive joints. Therefore, in case of changing the driven joint (for example from 1S1 to 2S1), the configuration shown in Figure 1b will be no longer singular. Geometric Algebra Approach to Singularity of Parallel Manipulators If we consider the blade formed only by the duals associated with the dummy joints, the condition for the so-called constraint singularity (the term introduced by Zlatanov et al., 2002) can be obtained, i.e., 1Dd \u2227 2Dd = 0 or 1D\u0303d \u2227 2D\u0303d = 0. (22) Therefore, the constraint singularity occurs when these two lines 1Cd = 1D\u0303d and 2Cd = 2D\u03032 coincide (Figure 2). From Eqs. (19) and (22) it can be seen that the constraint singularity is a subset of general singularity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003724_ccdc.2015.7162172-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003724_ccdc.2015.7162172-Figure2-1.png", + "caption": "Fig. 2 Frame system used in quadrotor dynamic modelling", + "texts": [ + " But states which describe the attitude of a aerial vehicle remain the same: the velocity corresponding to body axis (u, v, w), the angular rate toward body axis (p, q, r), the position toward earth axis (x, y, z), the Euler angle roll, pitch, yaw (\u03c6, \u03b8, \u03c8). As the sensors onboard are not sufficient for all information, only 6 states (p, q, r, \u03c6, \u03b8, \u03c8) are observable without external support. Though the dynamics of quadrotor is relatively simpler than helicopter, the analysis to determine the whole dynamics model is in great length. The way deriving this dynamic model will not be discussed in this paper. Only a brief conclusion sufficient to support section 4 will be given. The index of coordinates of quadrotor are defined in figure 2. There are two frames used in this dynamic analysis, the earth frame E and the body frame B. The earth frame is fixed to the earth ground. And the body frame is on the aircraft itself. In the earth frame, the resultant force of lift force and gravity decide the motion on z-axis. The kinematics and dynamics on z-axis is as:{ z\u0307 = w w\u0307 = (U1 \u00b7 cos\u03c6 \u00b7 cos \u03b8 \u2212m \u00b7 g)/m . (1) The relation between earth frame can be represented by a transformation. Here the details of transformation will not be discussed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001864_b978-0-12-382098-3.00006-8-Figure6.7-1.png", + "caption": "FIGURE 6.7 Rotating Couette flow.", + "texts": [ + " For 185 Chapter | 6 Rotating Cylinders, Annuli, and Spheres e>3:3pffiffiffiffiffiffiffiffiffi \u00f06:9\u00de \u03c4o =\u03c1 where e is the roughness grain size and \u03c4o is the shear stress at the surface, the moment coefficient is given by 1 a pffiffiffiffiffiffiffiffi \u00bc 1:501 \u00fe 1:250ln \u00f06:10\u00de Cmc e The center line average roughness, Ra, is related to the sand-grain roughness by Ra\u2248 ks/2 (see Childs and Noronha, 1997), allowing the following substitution. 1 a pffiffiffiffiffiffiffiffi \u00bc 0:8079 \u00fe 1:250ln \u00f06:11\u00de Cmc Ra The term Couette flow describes flow between two surfaces that are in close proximity, such that flow is dominated by viscous effects and inertial effects are negligible. In cylindrical coordinates this involves flow in an annulus, as illustrated in Figure 6.7, and the Navier-Stokes equations can be solved exactly by analytical techniques, subject to a number of significant assumptions, which severely limit the application of the resulting solution. Nevertheless, the approach is instructive and also forms the basis for a technique to determine the viscosity of fluid. The analysis presented here for rotating Couette flow assumes laminar flow and is valid provided the Taylor number given in a form based on the mean annulus radius in Equation 6.12, is less than the critical Taylor number, Tacr ", + " If the critical Taylor number is exceeded, toroidal vortices can be formed in the annulus. The formation of such vortices is considered in Section 6.4. The critical Taylor number is dependent on a number of factors, including the rotation ratio and annulus dimensions. For the case of a narrow gap annulus, with a stationary outer cylinder the critical Taylor number is 41.19. 1:50:5\u03a9r \u00f0b\u2212a\u00demTam \u00bc \u00f06:12\u00de where rm = (a + b)/2. \ufffd \ufffd \ufffd \ufffd \ufffd \ufffd 186 Rotating Flow If the flow is assumed to be contained between two infinite concentric cylinders, as illustrated in Figure 6.7, with either cylinder rotating at speed \u03a9a and \u03a9b, respectively, with no axial or radial flow under steady conditions, the continuity equation reduces to Equation 6.13 and the radial and tangential and components of the Navier-Stokes equations reduce to Equations 6.14 and 6.15. \u2202u \u00bc 0 \u00f06:13\u00de \u2202 2\u03c1u dp \u2212 \u00bc \u2212 \u00f06:14\u00de r dr d2u 1 du u 0 \u00bc \u03bc \u00fe \u2212 \u00f06:15\u00de dr2 r dr r2 or d2u d u \u00fe \u00bc 0 \u00f06:16\u00de dr2 dr r These equations can be solved with appropriate boundary conditions to give the velocity distribution as a function of radius and the torque on the inner and outer cylinders: The boundary conditions are as follows: at r = a, u = \u03a9a at r = b, u = \u03a9b, and p = pb Equation 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001801_robot.2010.5509380-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001801_robot.2010.5509380-Figure2-1.png", + "caption": "Fig. 2. Experimental setup used to perform 100 needle insertions.", + "texts": [ + " The covariance can be given as a closed form: \u03a3(t) = \u222b t 0 Ad\u22121 m(\u03c4)D0Ad\u2212T m(\u03c4)d\u03c4 = a1 0 0 0 a2 a3 0 a4 a5 a6 0 0 0 a5 a7 a8 0 0 0 a6 a8 a9 0 0 a2 0 0 0 a10 a11 a3 0 0 0 a11 a12 , (10) where a1 = (\u03ba2\u03bb2 2 + \u03bb2 3)t, a2 = \u2212\u03ba\u22122v\u22121\u03bb2 3(1 \u2212 C) a3 = \u03ba\u03bb2 2t \u2212 \u03ba\u22121\u03bb2 3((\u03bav)\u22121S \u2212 t) a4 = \u03bb2 1( 1 2 t \u2212 SC(2\u03bav)\u22121), a5 = \u03bb2 1S 2(2\u03bav)\u22121 a6 = \u03ba\u22122v\u22121\u03bb2 1(1 \u2212 C \u2212 1 2 S2), a7 = \u03bb2 1( 1 2 t + SC(2\u03bav)\u22121) a8 = \u03ba\u22121\u03bb2 1(S(\u03bav)\u22121 \u2212 1 2 t \u2212 SC(2\u03bav)\u22121) a9 = \u03ba\u22122\u03bb2 1( 3 2 t \u2212 2S(\u03bav)\u22121 + SC(2\u03bav)\u22121) a10 = \u03ba\u22122\u03bb2 3( 1 2 t \u2212 SC(2\u03bav)\u22121) a11 = \u03ba\u22123v\u22121\u03bb2 3( 1 2 S2 + C \u2212 1) a12 = \u03ba\u22122\u03bb2 3( 3 2 t \u2212 2S(\u03bav)\u22121 + SC(2\u03bav)\u22121) + \u03bb2 2t Here S and C denote sin(\u03bavt) and cos(\u03bavt), respectively. We used the device shown in Fig. 2 to perform repeated needle insertions. The needle was inserted into the artificial tissue by a DC servo motor attached to a linear slide. An additional DC motor rotates the needle shaft, but this was only used to orient the needle before insertion in these experiments. A stepper motor attached to the platform holding the artificial tissue was used to move the tissue between trials. We used a 0.57 mm diameter Nitinol wire (Nitinol Devices and Components, Fremont, CA, USA) with a bevel angle of roughly 45\u25e6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003725_icuas.2015.7152366-Figure9-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003725_icuas.2015.7152366-Figure9-1.png", + "caption": "Fig. 9. Wind tunnel acquisition system set-up.", + "texts": [], + "surrounding_texts": [ + "The mathematical model provides the means to exploit isolated aerodynamic and propulsion data and predict the behavior of the whole vehicle and reason over its design features (e.g., mass, airfoil, propeller geometry). Therefore, except for ki, the mathematical model parameters are either pure geometric, aerodynamic or propulsion quantities and can be identified separately. While geometric quantities can be directly measured, aerodynamic and propulsion coefficients call for wind tunnel measurements. For the sake of completeness, this section superficially describes the wind tunnel campaign that supported this work (see [11] for more information) and how the longitudinal subset of the collected data was used to interpolate the aerodynamic coefficients, tune the interaction factor ki and validate the model. Propulsion identification was carried out in [12] and the respective parameters (among all others identified in this work) can be found in the appendix section. The experiments were ran at the SabRe closed-loop wind tunnel (Fig. 6) located at ISAE and capable of delivering low Reynolds stable and uniform flow at a wind velocity range of 2 to 25 m/s, thus ideal for experimenting full-span micro air vehicles. Although a 6-component study was performed (3-dimensional forces and moments), this paper focus only on longitudinal quantities, i.e., drag D, lift L and pitching moment M measured with zero sideslip. Forces and moments were measured by means of a calibrated 5-component internal balance in two different configurations (see figure 7) in order to obtain the 6 force/moment components (see [11] for more information). An adapted MAVion was manufactured for wind tunnel campaign purposes (Fig. 1). Its main objectives were to enable rigid installment of the internal balance in both configurations and to provide a non-deformable airfoil section to aerodynamic identification. Fig. 8 illustrates the electronic counterpart. Notice that elevon deflections were measured by means of potentiometers installed directly at the elevons avoiding servo measurements that are susceptible to inaccuracy due to rod deformations and servo-control errors. On the other hand, propellers speed were measured by the brushless CC motors speed controllers. The campaign data set is roughly divided in two parts. The first was taken by independently setting the angle of attack \u03b1, motor speed \u03c9 and flap deflection \u03b4 to the values illustrated by table I. All combinations were exhaustively explored and the associated forces and moments registered. Notice that the unconventional nature of the vehicle calls for an unconventional envelope of testing variables that include very high angles of incidence. Fig. 10 isolates aerodynamic data from propulsion data by plotting the experimental aerodynamic coefficients in the wind coordinate system for \u03c9 = 0 and different elevon deflections. Each coefficient point is calculated by means of equations 5, 6 or 7, and later interpolated by least mean-squares to fit equations 9, 10 and 11. The results are shown in Fig. 10. The second data set comprises of wind relative velocity, elevon deflection and propeller rotation required to achieve static equilibrium in a given angle of attack. The experimental procedure fundamental idea is to, as previously done in [13], for each sampled \u03b1, iteratively search for values V\u221e, \u03b4i, \u03c9i that will deliver aerodynamic/propulsion forces and moments (notice gravitational suppression) that will cancel gravitational forces and moments on the flying model; which has a different mass distribution than the wind tunnel model that is not meant to fly. The desired flying model mass distribution is such that the center of mass is longitudinally located at a position 0.15c away from the leading edge and is motivated by a 10% stability margin controls requirement. The results can be seen in Fig. 11 along with power required curves. Furthermore, at this stage, all mathematical model parameters are identified except ki which is determined by ki = arg min ki Np\u2211 j=1 f(xj ,uj ,0)TSf(xj ,uj ,0) (41) where xj = ( vn,j 0 cos( \u03b1j 2 ) sin( \u03b1j 2 ) 0 )T (42) and uj = ( \u03b4j \u03c9j )T (43) where Np is the number of experimental points, {vn,j , \u03b4j , \u03c9j} are the experimental equilibrium variables for a given \u03b1j , and S is a diagonal positive definite scaling matrix to account for different orders of magnitude between f(\u00b7) components. Finally, after identifying ki, the mathematical equilibrium curves can be plotted by means of numerically solving f(x,u,0) = 0 and compared with experimental values for model validation. The results can be seen on Fig. 11. The results validate the design by demonstrating that the MAVion is capable of sustaining flight from 0m/s to 20m/s in the absence of wind. Equivalently speaking, the MAVion is capable of maintaining hover flight in adverse wind conditions up to 20m/s. It is noted that MAVion maximum speed was not reached in wind tunnel testing due to internal balance strain gauge saturation and it is still an unknown. However, the blow-up in propeller engine rotation \u03c9 for small angles of attack in the equilibrium figures suggests an upper bound of vn \u2264 20m/s. Finally, the elevator deflection angle shows a maximum of \u03b4 = 27o at \u03b1 = 40o, within the range of the elevon aerodynamic efficiency. This confirms that the propeller slipstream is strong enough to guarantee pitch control throughout the entire transition flight. Finally, the mathematical model allows for drag polar computation (see Fig. 14) which defines the necessary angle Require: Q,R > 0 for acceptable hover flight performance Require: \u03c1init > 0,\u2206\u03c1 > 0 i\u2190 0 (xi,ui)\u2190 (xh,uh) while \u2203p \u2208 F : p /\u2208 \u222aik=0R(\u03c1k) do i\u2190 i+ 1 \u03c1\u2190 \u03c1init A\u2190 \u2202f \u2202x \u2223\u2223\u2223 x=x0,u=u0 , B \u2190 \u2202f \u2202u \u2223\u2223\u2223 x=x0,u=u0 Compute LQR gain Ki for linear system x\u0307 = Ax\u0303+Bu\u0303 while x\u0307TP x\u0303+ x\u0303TP x\u0307 < 0 \u2200x\u0303 : V (x\u0303) < \u03c1+\u2206\u03c1 do \u03c1\u2190 \u03c1+ \u2206\u03c1 end while \u03c1i \u2190 \u03c1 Find (x\u2217,u\u2217) \u2208 F : (x\u2217 \u2212 xi) TP (x\u2217 \u2212 xi) = 0.70\u03c1i (xi+1,ui+1)\u2190 (x\u2217,u\u2217) end while Fig. 13. Summary of equilibrium trajectory fixed point sampling and global scheduled LQR controller generation algorithm. of attack \u03b1h (and consequently xh and uh) for maximum endurance. Such point will be defined as MAVion\u2019s horizontal flight cruise equilibrium point. On the other hand, the point where vn = 0, i.e. hover equilibrium points xv and uv , can be as well easily computed and both valors are illustrated at table II among other flight quantities that derives from the results from this section." + ] + }, + { + "image_filename": "designv11_3_0001843_1.39537-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001843_1.39537-Figure5-1.png", + "caption": "Fig. 5 Top and side view of the system after the virtual structure rotation.", + "texts": [ + " Because the principal axis, l, is normal to the vectors S and XT, the principal axis vector can be defined as l S XT jS XTj (23) Throughout the virtual structure rotation, each satellite moves to a new position, which can be calculated by using Eq. (22) as A 0 X XA0 X R2XA (24) B 0 X XB0 X R2XB (25) C 0 X XC0 X R2XC (26) Figure 4 shows the qualitative procedure of the virtual structure rotation. Note that S0 is a boresight vector after rotation, and axis l is normal to both S and XT in Fig. 4. C. Third Stage: Targeting The attitude of each satellite is coincided with the VSF in the first and second stages, as shown in Fig. 5. Figure 5 shows the top and side views of the system after the virtual structure rotation. Let us first consider satellite A. The satellite only needs to rotate at the rotationFig. 3 Virtual structure model in 3-D space. D ow nl oa de d by W R IG H T S T A T E U N IV E R SI T Y o n Se pt em be r 11 , 2 01 4 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /1 .3 95 37 angle A about the new principal axis lA to point toward the desired target. Because the principal axis lA is normal to the vectors z\u0302A0 and AX, lA can be determined" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001937_icems.2011.6073463-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001937_icems.2011.6073463-Figure4-1.png", + "caption": "Fig. 4. Schematic diagram of the 4-mass drive train modeling", + "texts": [ + " 3, the rigid blade (OA) and the flexible blade (AB) [7]. The blade sections OA1, OA2 and OA3 are collected into the hub and have the inertia HH, and the rest of the blade sections A1B1, A2B2 and A3B3 are simplified as a disk with small thickness with inertia HB about the shaft. It is assumed that the three-blade turbine has uniform weight distribution; and the turbine torque is equal to the summation of the torque acting on three blades. As a result, 4-mass model is developed from 6-mass and shown in Fig. 4. In addition, KBH is the stiffness of the blades, and DBH is the mutual-damping between the flexible blade disk and hub. And the equations of motion for 4-mass model can be expressed as 0 2 [ ( )] 2 [ ( )] [ ( )] 2 [ ( )] [ ( )] 2 [ ( )] ( B B W BH BH BH B H B B H H BH BH BH B H LS LS LS H GB H H GB GB LS LS LS H GB HS HS HS GB G GB GB G G HS HS HS GB G G G G BH B dH T K D D dt dH K D K D D dt dH K D K D D dt dH K D D T dt d dt \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 = \u2212 + \u2212 \u2212 = + \u2212 \u2212 + \u2212 \u2212 = + \u2212 \u2212 + \u2212 \u2212 = + \u2212 \u2212 \u2212 = \u2212 0 0 ) ( ) ( ) H LS H GB HS GB G d dt d dt \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u23a7 \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa\u23aa \u23a8 \u23aa \u23aa \u23aa \u23aa \u23aa = \u2212\u23aa \u23aa \u23aa = \u2212 \u23aa\u23a9 Where \u03c9B, \u03c9H, \u03c9GB, \u03c9G are the flexible blade disk, hub, gearbox and generator rotor rotational speed respectively; \u03c90 denotes the electrical system synchronous speed; \u03b8BH is the angle between the flexible blade disk and the hub; \u03b8LS is the (3) angle between the hub and the gearbox, \u03b8HS is the angle between the gearbox and the generator rotor. C. 3-Mass I & II Drive Train Modeling The 4-mass model can be converted into a 3-mass model by adding the masses of two disks together. As shown in Fig. 4, there are two ways to get 3-mass model. The 3-mass I model can be obtained by adding the blade inertia and hub inertia together (inertia: HBH); while the 3-mass II model can be acquired by adding the gearbox inertia and generator rotor inertia together (inertia: HGBG). Fig. 5(a) illustrates the 3-mass I model, where the blades and the hub are assumed as rigid connection. And in 3-mass II model as shown in Fig. 5(b), the flexibilities of the high-speed shaft are ignored. The equations of motion for 3-mass I model are given as 0 0 2 [ ( )] 2 [ ( )] [ ( )] 2 [ ( )] ( ) ( ) BH BH W LS LS LS BH GB BH BH GB GB LS LS LS BH GB HS HS HS GB G GB GB G G HS HS HS GB G G G G LS BH GB HS GB G dH T K D D dt dH K D K D D dt dH K D D T dt d dt d dt \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u03b8 \u03c9 \u03c9 \u03c9 \u23a7 = \u2212 + \u2212 \u2212\u23aa \u23aa \u23aa = + \u2212 \u2212 + \u2212 \u2212\u23aa \u23aa\u23aa = + \u2212 \u2212 \u2212\u23a8 \u23aa \u23aa = \u2212\u23aa \u23aa \u23aa = \u2212\u23aa\u23a9 Where \u03c9BH is the wind turbine rotational speed; DBH is the damping coefficients of the wind turbine" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000468_tmag.2007.915894-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000468_tmag.2007.915894-Figure1-1.png", + "caption": "Fig. 1. Configuration of one-phase TFLM.", + "texts": [ + " network (3-D EMCN) method [5] considering magnetic nonlinearity and 3-D magnetic path. The parameters such as flux linkage, thrust, attraction force, and core loss are calculated with the flux and flux density, and the parameters are used as a form of lookup table in the dynamic simulation model. Before using the parameter for the dynamic simulation, the parameters are verified by comparing with measured ones. Finally, in order to verify the usefulness of the method, calculated current by the dynamic simulation and measured current are compared for the example solid core TFLM. Fig. 1 shows the configuration of a permanent magnet (PM) type TFLM fabricated for a high-power transportation system application. Both the mover and stator have solid cores for rigid fabrication, and PM and armature coils are in the mover. The principle of force generation of objective model is shown in Fig. 2, which is the cross section of Fig. 1. In the mover poles, two magnetic polarities by PMs, and , are changed to one polarity or by offsetting one polarity against the polarity of current coil. Therefore, when the polarity of current coil is , ideally there is only polarity in the mover. Alternating current functions as a switch turning on and off the mover polarity, therefore, mover and stator generate the total thrust in one direction. 0018-9464/$25.00 \u00a9 2008 IEEE As shown in Fig. 2(a), the sine wave source makes better output characteristics; the output power is higher and the thrust ripple is lower if more than two phases are used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001108_j.saa.2010.03.012-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001108_j.saa.2010.03.012-Figure4-1.png", + "caption": "Fig. 4. Relative fluorescence intensity \u02db as a function of log[Ag+]. The curves fitting the experimental data were calculated from Eq. (6): (1) m:n = 3:1, K = 5.1\u00d71016; (2) m:n = 2:1, K = 1.4\u00d7108; (3) m:n = 1:1, K = 3.7\u00d7105; (4) m:n = 1:2, K = 3.7\u00d71010. ( ) Data points experimentally obtained.", + "texts": [ + " It can be exper- mentally determined by measuring the fluorescence intensity of 1 n the solution: = [A]f [A]t = F \u2212 F0 Fb \u2212 F0 (5) Here Fb is the fluorescence intensity of T(p-OCH3)PPH2 in the lank buffer solution and F0 represents the fluorescence intensity f T(p-OCH3)PPH2 in the solution when T(p-OCH3)PPH2 is comletely complexed with Ag(I). F is the fluorescence intensity of (p-OCH3)PPH2 actually measured when in contact with Ag(I) soluions of a given concentration. The relationship between the \u02db and g(I) concentration [B] can be represented as \u02dbn 1\u2212 \u02db = 1 nK[A]n\u22121[B]m (6) The response of T(p-OCH3)PPH2 for different concentrations of g(I) is shown in Fig. 4. Seven curves are calculated using Eq. (6) ith different K and ratios of Ag(I) and T(p-OCH3)PPH2. It can be een that the curve with 1:1 complex ratio and an appropriate K f 3.7\u00d7105 fits best to the experimental data. The curve can serve s the calibration curve for the detection of Ag(I) concentration. practically usable range for quantitative determination covers a ange from 1.0\u00d710\u22127 to 5.0\u00d710\u22125 M [20]. The detection limit is .5\u00d710\u22128 M. Thus, a possible structure of Ag(I)-T(p-OCH3)PPH2 complex ould be proposed, in which tetrapyrrolic center of porphyrin could oordinate with silver ion (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001549_c0jm01523g-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001549_c0jm01523g-Figure4-1.png", + "caption": "Fig. 4 A schematic diagram of the electrochemical cell setup. FDH is immobilized on carbon electrodes with FDH-(F127-MST).", + "texts": [ + " As you can see, the reduction of NAD+ proceeds in a coupling reaction with oxidation of formaldehyde in the FDH-(F127MST). That leads our next interest toward investigation of the electron transfer from NADH to electrode. Therefore, electrochemical experiments using FDH-(F127-MST) were conducted in order to examine whether electron transfer successfully proceeds in the presence of NAD+ and quinone as a mediator. The details of the electrochemical set-up are described in the experimental section and Fig. 4. The electron transfer from formaldehyde is based on the following sequence of reactions: HCHO + NAD+ + 3H2O \u2013FDH/ HCOO + NADH + 2H3O+ (2) H+ + NADH + Q / NAD+ + QH2 (3) Fig. 5 Typical responses of (A) FDH\u2013(F127-MST) (1.7 mg: 0.01 mg as FDH) and (B) native (0.02 mg as FDH) to the injections of a formaldehyde solutions (1.4 and 14 ppm) at a constant potential of Eapp\u00bc 0.35 V. The electrolytes comprise 0.63 mM NAD+ and 0.64 mM quinone in a 12 ml of phosphate buffer (pH 7.4). 254 | J. Mater. Chem., 2011, 21, 251\u2013256 QH2 \u2013electrode/ Q + 2e + 2H+ (4) where FDH catalyses the oxidation of formaldehyde to formic acid, while NAD+ is reduced to NADH" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002294_iros.2011.6094627-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002294_iros.2011.6094627-Figure1-1.png", + "caption": "Fig. 1. Modified McGeer running model with point foot and hip torque T . The swing leg rotates with retraction rate \u03c9 and the foot has tangential speed vt immediately prior to contact with the ground. After contact, the foot can slip due to finite friction. When the foot does come to rest with respect to the ground, it acts as a pivot point until the leg spring free length l0 is reached and takeoff occurs.", + "texts": [ + " The running model proposed by McGeer in [11] has legs with distributed mass that can yield non-zero impact losses, which makes the model suitable for analyzing energy efficiency. In this study, we use a variation of the McGeer running model consisting of a point mass body with two rigid legs of distributed mass, each connected by a massless spring to a massless point foot. A torque actuator between the legs provides energy and control as the model runs over a flat, level surface. The three distinct phases of motion of the model are shown in Figure 1: flight, touchdown, and stance. The equations of motion for each phase are derived using Lagrangian mechanics. During flight and touchdown there is relative velocity between the front foot and the ground, so x, y, \u03b81, 978-1-61284-456-5/11/$26.00 \u00a92011 IEEE 3957 and \u03b82 are used as the generalized coordinates. While both of these phases have four degrees of freedom, interaction between the front foot and the ground must be modeled during touchdown with generalized forces. Friction is assumed to be Coulombic while there is relative motion between the front foot and the ground; collision of the other foot with the ground is ignored for simplicity", + " Using the previous solution as a seed, the gait resulting in the locally minimal objective function value is found for each value of \u03c9. The procedure is performed using the parameters of Table I and repeated for several body apex horizontal speeds, body apex heights, and leg masses. Example optimal torque profiles for several swing leg retraction rates are shown in Figure 3. Figure 4 shows energy expenditures and the corresponding foot tangential speed vt (magnitude of the velocity of the foot in the direction perpendicular to the leg at the instant of touchdown, as illustrated in Figure 1) as functions of \u03c9 for the parameters in Table I. The minimum of the touchdown loss curve and the zero of the foot tangential speed curve occur at the same retraction rate, as anticipated by analysis presented in [10]. What is significant and non-intuitive is that the total work required to drive the legs Eloss,swing corresponds closely with the touchdown loss. As a result, the minimum of the cmt curve and the zero of the foot tangential speed curve occur at nearly the same retraction rate. This is typical for all parameters sets studied" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000546_detc2009-86947-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000546_detc2009-86947-Figure1-1.png", + "caption": "Figure 1. GEOMETRICAL DIMENSIONS OF THE BATAVUS BROWSER BICYCLE SHOWN WITH DATA AQUISITION EQUIPMENT.", + "texts": [ + " Estimates of these properties can be determined with a detailed CAD model but we chose to measure the quantities for accuracy and time considerations. The bicycle was assumed to be made up of four rigid bod- 1 Copyright c\u00a9 2009 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/31/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use ies: the rear frame (B f ), the front wheel (F), the rear wheel (R) and the handlebar/fork assembly (H). Fifteen geometrical measurements (Fig. 1) of the bicycle were taken using a ruler (\u00b10.002 m) and an angle gage (\u00b10.5 deg). Only five of the measurements, w, c, \u03bb 1, rR and rF, are required for the benchmark model (Tab. 12). The rest of the measurements are used to estimate the seated position of the rider described in the HUMAN PARAMETER ESTIMATION section. We use the same global coordinate system as the benchmark model. The origin is at the rear wheel contact point with the X-axis pointing forward along the ground, the Z-axis downward and the Y -axis to the right (Fig. 1). All of the dimensions were taken as if they were projections into the XZ-plane except for the hub widths2. Note that in the model the top tube is assumed to be horizontal and the measurements were taken from the intersections of tube centerlines. The wheel radii were measured by rolling the bicycle forward with the rider seated on the bicycle for nine revolutions of the wheel. The distance traversed along the ground was measured with a ruler, divided by nine and converted to wheel radii using the relationship between radius and circumference, r = c 2\u03c0 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000871_09544062jmes949-Figure13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000871_09544062jmes949-Figure13-1.png", + "caption": "Fig. 13 Transition of geometry of FS between its cylindrical and toothed portions (FEM and actual HD)", + "texts": [ + " Therefore, it can be assumed that the implementation of such a fine grid to the teeth of the FS would have been beneficial by providing a smoother stress gradient within the teeth and a better estimate of the torsional stiffness of the HD. However, this refined model was not adopted for the FEM of the HD, as it would have resulted in a significant increase in the number of elements and of CPU time required to obtain the results. On actual HD, the FS is machined to join the teeth smoothly to the thin part of its cup. In the FEM, this smooth transition between the teeth and the cylindrical part of the FS is not modelled because the brick elements are joined to the shell elements using multipoint constraint (MPC) equations (Fig. 13). This modelling approach reduces the number of elements of the FEM, but it also slightly reduces the stiffness of the FS and causes significant stress levels around the junction of the solid and shell elements. This should therefore be considered if the currentmodel is used toperformstress analyses at this location of the FS. Apart from this local loss of accuracy due to the MPC, the FEM of the FS is properly meshed and provides realistic stress levels [6,11] and gradients. Using adesktop computerwith dual core processors clocked at 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001028_s11424-009-9157-7-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001028_s11424-009-9157-7-Figure1-1.png", + "caption": "Figure 1 3DOF helicopter", + "texts": [ + " In this paper, the helicopter model used in the experiment is a laboratory-scale 3DOF helicopter produced by Quanser Consulting, Inc. The 3DOF helicopter control system is a nonlinear MIMO uncertain system with unknown constant parameters, bounded disturbance, and nonlinear uncertainty. Our control goal is to have the attitude of the helicopter track a reference signal by output feedback. This situation corresponds, for instance, to the case in which a miniature unmanned helicopter has space limit and load limit for sensors, so only output is available for control scheme. The 3DOF helicopter plant is shown in Figure 1. It has two DC motors, which are mounted at the two ends of a rectangular frame and drive two propellers. The helicopter frame is suspended from the end of a long arm and is free to pitch about its center. A positive voltage to either motor also causes an elevation of the body. If the body pitches, the thrust vectors Yao YU Department of Automation, Tsinghua University, Beijing 100084, China. Email: yu-y04@mails.tsinghua.edu.cn. Yisheng ZHONG Tsinghua National Laboratory for Information Science and Technology, Department of Automation, Tsinghua University, Beijing 100084, China" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000240_0076-6879(81)76134-2-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000240_0076-6879(81)76134-2-Figure4-1.png", + "caption": "FIG. 4. Dilution valve operation. (A) Stopcock bore turned to allow flushing of both bell", + "texts": [ + " I~t l l I IN-~ ALUMINUM COVER PLATE IIIII:::::::IHII THIN LAYER CELL PLATE I- - -q~:~l t l \" / I~) REACTION BRASS VALVE '~II~IY/ ' , /7//M II_~ C H A M B E R HOUSING f l i ~ C O P P E R VISE STAIN LESS ST EEL J:,!i . ~ ~ / / Z I \"aH- ' t~PLATE VALVE ' PC~PP~I~ C : H p p ~ l ~ T ~ ' ~ . . . . . . . . . i~k~COPPER THERMAL ~ '~+ ~ . . . . . . . t_k~ CONTROL TUBING ~ o O-RINGS Fl\u2022. 3. Spectrometer cell holder and dilution valve assembly. With the apparatus assembled and placed in the spectrophotometer, one can turn the dilution valve to a position shown in Fig. 4, where the bore of the valve may be flushed with the known incoming gas (typically oxygen) and where the bore connects to the optical cell. The spectra of the hemoglobin sample may be monitored to assure the stability and condition of the layer. In some instances, to assure that the cell is filled with incoming gas one may wish alternately to flush bore and turn to the cell connected position, wait 30 sec for gas diffusion to be completed, and repeat the procedure 10-20 times. This has the advantage that the thin layer is exposed only to atmospheric pressures, since an intermediate position of the bore (Fig. 4C) sets the pressure to atmospheric conditions. Oxygen Partial Pressure Preparation and Measurement The precision dilution valve serves to create a set of partial pressures that are related to each other by the dilution ratio, Q, of the \"cel l\" volume to the total volume of \"cel l\" plus valve bore. Ifpo is the original oxygen partial pressure within the cell and flushed bore, and if one then flushes the bore with nitrogen and turns the bore to connect with the cell (Fig. 4B,C,D), then the partial pressure of oxygen within the system drops by the dilution factor to p l = Qpo (1) and bore with flowing gas. (B) Bore flush position. (C) Equilibration of gas pressure to atmosphere position. (D) Bore to cell position. Normal step operation involves steps B, C, D (clockwise, followed by counterclockwise rotation). Repeating this process produces another drop to give P~ = Q p l o r P2 = Q2po (2) We can then see that the partial pressure achieved at the end of the ith step is P, = Q~po (3) IXo, = iX\u00b0o, + R T In Po~ (4) where ~\u00b0o, is the standard state chemical potential, R the gas constant, and T the absolute temperature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003855_j.triboint.2015.03.025-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003855_j.triboint.2015.03.025-Figure1-1.png", + "caption": "Fig. 1. Geometrical domain of the EHL part (left) and overall geometry of the contact with associated heat transfer mechanisms (right).", + "texts": [ + " The numerical model employed in this work is based on the full-system finite element approach developed in recent years by the authors and described in detail in [15,16]. Here, only its main features are recalled. In Section 3.1, the EHL part is described and its corresponding equations recalled. The thermal part is described in Section 3.2 and finally, the overall numerical procedure is recalled in Section 3.3. The EHL part of the model consists of three main equations: the generalized Reynolds applied to the 2D contact area \u03a9c(see Fig. 1 (a)); the linear elasticity equations applied to the 3D solid domain and the load balance equation. The Reynolds equation for a steadystate point contact between a sphere and a flat plane lubricated with a generalized Newtonian lubricant under unidirectional surface velocities in the x-direction is given by Yang and Wen [23] \u2202 \u2202x \u03c1 \u03b7 e h3 \u2202p \u2202x \u00fe \u2202 \u2202y \u03c1 \u03b7 e h3 \u2202p \u2202y \u00bc 12 \u2202 \u2202x \u03c1nUmh \u00f07\u00de #where:Um \u00bc up\u00feus 2 \u03c1 \u03b7 e \u00bc 12 \u03b7e\u03c10 e \u03b70e \u03c1\u2033e \u03c1n \u00bc \u03c10 e\u03b7e us up \u00fe\u03c1eup Um \u03c1e \u00bc 1 h Z h 0 \u03c1 dz \u03c10 e \u00bc 1 h2 Z h 0 \u03c1 Z z 0 dz0 \u03b7 dz \u03c1\u2033e \u00bc 1 h3 Z h 0 \u03c1 Z z 0 z0dz0 \u03b7 dz 1 \u03b7e \u00bc 1 h Z h 0 dz \u03b7 1 \u03b70e \u00bc 1 h2 Z h 0 z dz \u03b7 Indices p and s correspond to the plane and the sphere respectively and the film thickness h in Eq", + " Note that, under the range of operating conditions considered in this work, thin-film assumptions and the use of the generalized Reynolds equation are entirely justified and pressure variations across the film thickness should be negligible. Only the velocity profile might be inaccurate in the inlet region of the contact, as reported in CFD simulations of EHL contacts (see for instance [24]). But this has no influence on compressive heating. The normal elastic deformation of the solid surfaces is obtained by solving the linear elasticity equations on a large 3D solid body representing a half-space domain as shown in Fig. 1(a). Finally, the load balance equation is used to complete the EHL part of the modelZ \u03a9c p d\u03a9\u00bc F \u00f09\u00de This equation is used to ensure that the correct external load F is applied to the contact by monitoring the value of the rigid body displacement h0. The generalized Reynolds equation defined above is associated with the usual EHL boundary conditions. That is, zero pressure is assumed on the boundary of the contact area \u03a9cand the free boundary problem arising at the exit of the contact is handled by applying a penalty method as proposed by Wu [25]", + " As for the linear elasticity part, the pressure distribution obtained from the Reynolds equation is used as a normal pressure load boundary condition on the contact surface \u03a9c while a zero displacement boundary condition is applied on the bottom side \u2202\u03a9b of the 3D geometrical domain \u03a9 at a sufficient distance from the contact area to approximate a half-space configuration. Finally, a zero normal stress boundary condition is applied to the rest of the boundaries of\u03a9. The symmetry of the problemwith respect to the xz-plane is taken into consideration to reduce its associated computational cost. For more details the reader is referred to [16]. Heat is generated within the lubricant film by compression and shear as indicated earlier. Parts of the generated heat (Qs and Qp) are dissipated towards the solids by conduction as shown in Fig. 1(b). The solids then carry part of this received heat towards the contact exit by advection. Another part of the generated heat (QL) is carried out by the lubricant towards the exit of the contact by advection while some of the heat remains confined within the lubricant film leading to a local increase in its temperature. Heat transfer by conduction is dictated by the thermal conductivities of the solids and lubricant whereas advection is governed by their volumetric heat capacities and speeds. Temperature distribution in the two solid bodies and the lubricant film is obtained by solving the 3D energy equation applied to their respective geometrical domains" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002494_c3ra46562d-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002494_c3ra46562d-Figure1-1.png", + "caption": "Fig. 1 The parallel co-flowing coaxial microfluidic device used in the experiments. Red and green flows represent two inner-phase flows, they could be either silicon oil, deionized water with surfactant, tetradecane or gas, red flow and green flow contain two insoluble phases. Yellow and blue flows represent middle and outer phase flows. Yellow flow could be either TPGDA, HDDA or silicon oil. Blue flow is deionized water with 2% SDS and 1% PVA. Inner-phase flows are pumped into microfluidic device through two micro-needles, middle phase flow encapsulates them to form a first structure red & green droplets/ yellow flow, then outer flow engulfs the structure, so a red & green droplets/yellow droplet double emulsion is formed. To demonstrate the microfluidic device clearly, we make a cross section in the droplet forming area.", + "texts": [ + " Based on it, the physics of the structure evolution of double emulsions with two-phase cores allows the synthesis of microparticles with new morphologies and more applications in the encapsulation of multiple active substances. Two cross junctions are fabricated on a 40 mm 25 mm 7 mm polymethyl methacrylate (PMMA) chip using an end mill. The two inner phase uids are introduced trough the parallel micro-needles with inner-diameter of 160 mm and outer-diameter of 300 mm. The channel for middle and outer phase uids is approximately 1.50 mm wide 1.50 mm high, as shown in Fig. 1. A circular glass capillary with 1.05 mm inner-diameter and 1.5 mm outer-diameter is tapered using a micropipette puller (P-97, SUTTER Co. Ltd., USA) for the injection of the middle phase uid. The diameter of the tapered orice is approximately 340 mm. The rst tapered capillary is inserted into a second capillary for the outer phase and coaxality is guaranteed by matching the outer diameter of the capillary with the inner dimension of the channel. The second capillary is not tapered. Six PTFE pipes are inserted into sides of the chip channel for carrying middle and outer phase uids" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001117_045104-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001117_045104-Figure4-1.png", + "caption": "Figure 4. Photo of a ball artefact.", + "texts": [ + " Inspection of the gear checker using the arbitraryshaped artefact. The involute helicoid has a complex curved shape; therefore, it is difficult to manufacture an involute helicoid accurately, and its accuracy is estimated at approximately 1 \u03bcm. In contrast, a simple shape such as a ball or plane, can be manufactured very accurately. For example, a ball can be manufactured at the form accuracy of tens of nm. Therefore, the inspection method of the gear checker using the ball artefact as shown in figure 4, which consists of very precisely manufactured balls and plane, has been studied. In the case of calibrating the gear checker using such a noninvolute helicoid artefact, the theoretical measurement results of the artefact, in other words, the calculated measurement results assuming the use of a gear checker without any error factor and the actual measurement result using a real gear checker with some error factors are compared. The difference between them is taken as the calibration result. The theoretical measurement result varies according to the form checking ball radius (cf figure 4) and the distance between the balls of the ball artefact. The VGC enables the virtual measurement of an arbitrary-shaped object such as a ball artefact, and the theoretical measurement result can be calculated easily. The bold line in figure 5 shows the base circle (radius rb) and the corresponding involute curve. The involute curve and line AB (line of action) intersect at point T. Assume that the base circle and the involute curve rotate as one body by \u03b8 and then go to the position indicated by the dotted line in figure 5", + " It leads to the result that the uncertainty for the total profile or helix deviation has a value close to that for the profile and helix slope deviation. This result indicates that the uncertainty assessment using the VGC is possible. Figure 14 shows the theoretical measurement result of the ball artefact calculated using a VGC as an example. Assume that the base circle radius input into the gear checker is 40 mm, and the measuring probe radius is 1 mm. Figure 14(a) shows the theoretical measurement results, where the form checking ball radii (cf figure 4) are 9, 10 and 11 mm and the distance between balls (cf figure 4) is constant (40.14 mm). Figure 14(b) shows the results, where distances between balls are 40.04, 40.14 and 40.24 mm, and the form checking ball radius is constant (10 mm). The VGC facilitates the calculation of the theoretical measurement result of the ball artefact. To calibrate the gear checker using the ball artefact, two peaks of the camelback curve of the theoretical measurement result should have identical height [16, 17], and the geometry of the ball artefact should be designed to realize it" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000171_iros.2008.4650592-Figure5-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000171_iros.2008.4650592-Figure5-1.png", + "caption": "Fig. 5. Minimum and maximum path curvatures intersecting with a hazard.", + "texts": [ + " A potential field corresponding to the current desired waypoint location can then be defined as follows: 2)()( dgg KPF \u03ba\u03ba\u03ba \u2212= (11) where \u03bad is the desired steering angle and Kg is a positive gain parameter to modulate the potential field amplitude. A potential field for the desired UGV velocity can be simply expressed as follows: 2 1 )()( Kv dvv vvKvPF \u2212= (12) where vd is the desired velocity and Kv1 and Kv2 are positive gain parameters to modulate the potential field amplitude. vd can be a function of position to reflect mission objectives. Consider a UGV approaching a hazard as shown in Fig. 5. Here \u03ba1 and \u03ba2 are the maximum and minimum path curvatures that intersect the hazard. A potential field for hazard locations is constructed based on the following observations: - Path curvatures between \u03ba1 and \u03ba2 can be safely followed until the UGV is near the hazard; - The potential field magnitude should be greater at high speed than at low speed since control accuracy generally decreases with increasing speed; - Relative locations of waypoints and hazards should influence the hazard potential field value (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000744_1.4000484-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000744_1.4000484-Figure2-1.png", + "caption": "Fig. 2 Design results: \u201ea\u2026", + "texts": [ + " ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/pdfaccess.ash 4.1 The Sealing and Stress Region Analysis. This paper uses curvature difference analysis to estimate the sealing property. Let the trochoid ratio parameter =r /cN. Three cases are provided to illustrate how the sealing changes for different values of . If the parameters are N=6, r=30 mm, and R=5 mm and the values of are 2, 1.5, and 1.1, respectively, then the gerotor profiles are those shown in Fig. 2. Figure 3 shows the result of the curvature difference if the curvature difference is defined as d = i\u2212 0. In Fig. 3, for =2, the variation in curvature difference is from approximately 95\u2013140 deg. For =1.5, the angle range is about 110\u2013145 deg. For =1.1, the angle range is from 135 deg to 160 deg. This shows the result is the larger variation region of curvature that occurs near 180 deg side, and the angle range would be smaller when the parameter decreases. Figure 4 presents the rotor profiles, which show that the influence area and the parameter are directly proportional" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003810_j.isatra.2014.05.013-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003810_j.isatra.2014.05.013-Figure3-1.png", + "caption": "Fig. 3. Cart\u2013pole system.", + "texts": [ + " It is clear from (8) that x2-0 if and only if z-0 which implies that S2-0. Therefore, it can be concluded that the stabilization of both subsystems can be achieved. In order to verify the theoretical considerations and show the correct operation of the proposed control method, a cart\u2013pole system is simulated and comparisons between the proposed method and the existing decoupled methods (DSMC and TVSSS) are demonstrated. All simulations were carried out by Matlab/ Simulink. The dynamic behavior of the cart\u2013pole system shown in Fig. 3 can be described by the following nonlinear equations: _x1\u00f0t\u00de \u00bc x2\u00f0t\u00de _x2\u00f0t\u00de \u00bc f 1\u00f0x; t\u00de\u00feb1\u00f0x; t\u00deu\u00f0t\u00de\u00fed1\u00f0t\u00de _x3\u00f0t\u00de \u00bc x4\u00f0t\u00de _x4\u00f0t\u00de \u00bc f 2\u00f0x; t\u00de\u00feb2\u00f0x; t\u00deu\u00f0t\u00de\u00fed2\u00f0t\u00de \u00f035\u00de where f 1\u00f0x; t\u00de \u00bc mtg sin \u00f0x1\u00de mpL sin \u00f0x1\u00de cos \u00f0x1\u00dex22 L\u00f0\u00f04=3\u00demt mp cos 2\u00f0x1\u00de\u00de b1\u00f0x; t\u00de \u00bc cos \u00f0x1\u00de L\u00f0\u00f04=3\u00demt mp cos 2\u00f0x1\u00de\u00de f 2\u00f0x; t\u00de \u00bc \u00f04=3\u00dempLx22 sin \u00f0x1\u00de\u00fempg sin \u00f0x1\u00de cos \u00f0x1\u00de \u00f04=3\u00demt mp cos 2\u00f0x1\u00de b2\u00f0x; t\u00de \u00bc 4 3\u00f0\u00f04=3\u00demt mp cos 2\u00f0x1\u00de\u00de \u00f036\u00de where x1\u00f0t\u00de is the angular position of the pole from the vertical axis, x2\u00f0t\u00de is the angular velocity of the pole with respect to the vertical axis, x3\u00f0t\u00de is the position of the cart, x4\u00f0t\u00de is the velocity of the cart, mt is the total mass of the system (which includes the mass of the pole, mp, and the mass of the cart, mc), and L is the half-length of the pole" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001974_09544097jrrt341-Figure13-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001974_09544097jrrt341-Figure13-1.png", + "caption": "Fig. 13 FE mesh of the wheel\u2013rail contact: (a) showing the orientation of the track coordinate system, with X in the direction of rolling, (b) three-dimensional finite-element mesh of wheel\u2013rail, and (c) enlarged view of local meshed near contact point", + "texts": [ + " In the wheel\u2013rail contact property, static and kinetic friction and stickslip motion are described by friction model, in which friction coefficient is related with the interface slip rate. According to the article [17] in which the brake test was done in Lyon when wheels of vehicle were slipping on rails, for dry wheel\u2013rail contact surface, relation between friction coefficient and slip velocity is shown in Fig. 12, which is used in the simulation. The wheel\u2013rail rolling globe coordinate system is right-handed Cartesian, with \u2018x\u2019 being parallel to the tracks in the direction of rolling, \u2018y\u2019 horizontal to the right, and \u2018z\u2019 vertically down. The coordinate system is shown in Fig. 13(a). The right wheel and right track are selected to be analysed, and they are meshed by 3D finite-element meshes. The minimum size of the finite-element meshes is 1 mm near the wheel\u2013rail contact point and the model consists of 267 649 elements and 314 077 nodes, as shown in Fig. 13(b). The parts close to the contact surfaces of wheel\u2013rail are meshed more finely. The local meshes near contact surfaces are magnified and shown in Fig. 13(c). The wheel\u2013rail rolling contact finite-element model based on ALE FEM is developed and applied to the wheel\u2013rail steady-state rolling contact analysis. The von Mises stresses for static state contact of the wheel\u2013 rail are shown in Fig. 14, and those for steady-state contact of the wheel\u2013rail are shown in Fig. 15. It is shown in Fig. 14 that on no tangential traction, the maximum of the von Mises stresses is 642.6 MPa, which appears inside the bodies at 3.2 mm below the surface of the wheel. It is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000031_6.2008-7413-Figure6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000031_6.2008-7413-Figure6-1.png", + "caption": "Figure 6. Rotor bottom view (left), rotor top view (middle), and brushless motor (right).", + "texts": [ + " Each rotor has three CF rotor blades attached, where two opposing rotors are equipped with clockwise and the other two with counter-clockwise spinning blades. Rather than using toothed gears, the transmission between motors and rotors is performed via belt drives, which tend to be less noisy and are more robust for this purpose. The belt-drive gears and belts are industrial DIN-norm T2.5 components, which can be found in tools of any description, for instance, small planing machines. The rotor is composed of an AL belt-drive gear with press-fitted stainless-steel bearings at both ends, shown in Figure 6 (left). This main gear directly carries the three AL shafts for mounting the blades. The shafts are conjuncted by a TI triangle at the upper end, depicted in Figure 6 (middle), for compensating centrifugal forces that might cause shaft deformations at high rotation speeds. The rotor axles are also manufactured from TI to achieve high integrity at low weight. Note that the rotor axles not only carry the rotors, they also serve as primary frame components that connect the upper and lower CF pipes of the side arms. Compared to conventional quadrotor designs where only one end of a rotor axle is connected with a side arm, the present double-end-mounted rotor-axle design has two significant advantages: (1), it enables precise adjusting of the four rotor axles regarding plane parallelism, and (2), it prevents the rotors from undesired twisting that might occur with one-end-mounted rotors. In order to generate maximum thrust at minimum weight, we decided to use custom-built 3-phase AC motors also known as brushless motors or outrunners, shown in Figure 6 (right). They consist of CNCfabricated AL parts and hand-coiled armatures, are equipped with a TI axle, and weigh only 35g. They were originally designed for model planes, but re-shaped for us according to our specifications. The modifications applied are as follows: (1), the armature side was re-shaped to fit the JAviator\u2019s motor mount and all redundant material removed to enhance cooling. (2), the axle was placed upside-down, so that the axle\u2019s stub is located on the mounting side rather than the outrunner side" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001946_j.engfracmech.2011.07.012-Figure7-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001946_j.engfracmech.2011.07.012-Figure7-1.png", + "caption": "Fig. 7. (a) FE results for contact pressure pFE and (b) surface stress rx in the vicinity of the asperity for xd = 260.6 lm. The white semi-circle represents the asperity size and the position max(rx(x,xd)) is indicated with .", + "texts": [ + " 6, where a moving average filter was applied to reduce variations between consecutive positions of the cylindrical contact. During over-rolling p0p and q0p remained fairly constant, which through Eq. (1b) yielded a rather constant ap 6 22.5 lm. For the case study of Fig. 1a, the extreme values of the maximum pressure and traction were 5.0 GPa and 2.1 GPa, respectively. Outside the interval 363.2 lm < xd < 363.6 lm the asperity load was zero and the fatigue crack was only subjected to the cylindrical contact load. Fig. 7a and b shows pFE and the surface stress rx in the vicinity of the asperity for the cylindrical contact position that yielded max(rx(x,xd)), i.e. the maximum FE surface stress rx, for all xd and positions x along the symmetry line (y = 0). The results show that max(rx(x,xd)) occured near the border of the asperity contact. Observe also that a large zone in front of the asperity was not subjected to any contact pressure and experienced tensile surface stress rx. Fig. 7a illustrates that the asperity pressure was spherical with a circular radius ap. The FE surface stress rx(x,xd) was extracted along the symmetry line. Fig. 8a presents max\u00f0rx\u00f0x\u00de\u00dejxd , i.e. the maximum surface stress rx for varying x and fixed xd. Note that the tensile surface stresses induced by the asperity load were the highest when the asperity entered the cylindrical contact. Hence, max(rx(x,xd)) = 2.1 GPa and occured for xd = 260.6 lm, see Fig. 7b, and was located at x = 28.1 lm < rasp. Fig. 8b shows max(rx(xd))jx, i.e. the maximum surface stress rx for varying xd and fixed x. The FE geometry was not an infinite body, which explained non-zero rx when the asperity was not loaded by the cylinder. The zone with high tensile surface stress in front of the asperity, i.e. when x > 0, was identified as a potential RCF initiation region. An initial crack perpendicular to the surface would be subjected to rx. The RCF crack path was modelled in the symmetry plane (y = 0)", + " The real asperity shape may however differ in details from the cosine shape and therefore the real asperity pressure may differ in details from the spherical Hertz profile. Nonetheless, the major hypothesis remains, local asperities break through the lubrication film and create point loads with a local convex curvature that gives a tensile surface stress rx at contact entry. The initial crack was short, perpendicular to the surface and positioned at the maximum surface stress in front of the asperity. As long as the initial crack remained within the region with high surface stress in Fig. 7b, the simulated path was relatively unsensitive to xc. Other positions in this region would result in slightly shallower paths. Given the short size of the crack, the initial crack length and direction have low influence, while the crack will immediately kink to a profile quasi-identical to the one in Fig. 12a. The important result from the FE simulation was the distribution between the cylindrical and the asperity pressures including the sequence for them during the over-rolling. The continued analysis utilized that the pressure distribution at asperity over-rolling could be well approximated with a cylindrical Hertz pressure with superposition of a spherical Hertz pressure for the asperity contact, see Figs", + " 11d small differences were noted, primarily between the maximum tangential principal stress criterion, CPPC 1, and the other criteria. The KII differences were however judged as negligible but the oscillations that sometimes appeared for the four SIF based criteria were a clear drawback for these. Thus, the maximum tangential principal stress criterion was preferred. It gave accurate and stable solution while at the same time being straightforward to evaluate. The FE simulation with data from the gear contact in Fig. 1a resulted in a large tensile surface stress in front of the asperity, see Fig. 7b. The maximum von Mises stress in front of the asperity was approximately 3 GPa, which is higher than the monotonic yield stress in Table 2. However, the analyses focused on the steady-state solution after rapid isotropic hardening [10] to rY,cycl 1500 MPa in Table 2 and a kinematic shift in the stress space. Hence, the elastic stress range should be compared to the diameter of the cyclic yield surface 3 GPa. The high stress was limited to a small region, see rN in Fig. 12b. Initial yielding in an elastic\u2013plastic model would be limited to the absolute surface region and was not expected to affect the substrate compliance or contact pressures" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002891_iros.2011.6094953-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002891_iros.2011.6094953-Figure1-1.png", + "caption": "Fig. 1. 3D scan data (left) are reconstructed into a 3D ultrasound image (right).", + "texts": [ + " Since the 4D US scanner is designed for the purpose of research works, we can access digital data before the conversion into an image. Afterwards, a set of volume data is converted into a volumetric image, called as 3D scan conversion. For scan conversion, geometries of the probe (the 2D US transducer\u2019s radius Rprobe and the motor\u2019s radius Rmotor) and imaging parameters (the number of sample data in a A-line Nsamples, the number of A-lines in a frame Nlines, and the number of frames in a scan volume N f rames, angle between neighboring A-lines \u03b1line, and angle between neighboring frames \u03b1 f rame) are considered in (1). In Fig. 1, a sample s(i, j,k), which is the i-th datum along the j-th A-line in the k-th frame, is relocated into a point p(r,\u03d5,\u03b8) in a volumetric image, which is represented as p(x,y,z) in Cartesian coordinates according to (1) and (2). Note that our 3D US probe continuously scans volumes while its motor is sweeping the volume in forward and backward directions. Additionally, sweeping direction d (which is 1 in the forward direction and 0 in the backward direction in (1c)) and motor\u2019s rotating motion should be considered" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001025_j.snb.2008.07.017-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001025_j.snb.2008.07.017-Figure1-1.png", + "caption": "Fig. 1. A schematic of the sensor design.", + "texts": [ + " Recently there have een two useful advances in this field: firstly there has been the d w c g s o s c s t P c l p q o t 2 2 w S w B 5 C D E 2 P r t t t s w m i P o s c w c e x d T l r f r a 9 o g t w 3 3 c c T m o T i a t t a c g t g s a p m evelopment of a colorimetric sensor, whereby two lumophores ith different emission colours are integrated into a device and a olour change is seen in response to the partial pressure of oxyen [17,30], and secondly LEDs have become common excitation ources for sensors [18\u201323]. In this paper we describe a new type f sensor which integrates both of these technologies to give a ensor which is simple, compact and cheap, and, since it gives a olour response, it is easy to use with little training required. In this ensor a green LED is used as both a green emitter and the excitaion source for the oxygen-sensitive layer which is the red-emitting t octaethylporphyrin (PtOEP) dissolved in an ethyl cellulose film ontaining a scattering agent (Fig. 1). This device gives a \u201ctraffic ight\u201d, green\u2013yellow\u2013red, response on decreasing oxygen partial ressure; at high oxygen partial pressures emission from PtOEP is uenched and the green of the LED is the prominent colour; at low xygen partial pressures emission from PtOEP is unquenched, and he resulting red emission predominates [24]. . Experimental .1. Materials Ethyl cellulose (46% ethoxyl content), ZnO, and Degussa TiO2 ere purchased from Aldrich. PtOEP was obtained from Fischer cientific. \u201cSuperbright\u201d 5-mm LEDs (catalogue number: N62AX) ere bought from Maplin Electronics, Valley Road, Wombwell, arnsley S73 0BS" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003923_jpsj.83.043001-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003923_jpsj.83.043001-Figure1-1.png", + "caption": "Fig. 1. Experimental setup. (a) Annular container illuminated from the bottom. (b) Top view of the container. Spatio-temporal brightness distribution l\u00f0 ; t\u00de is recorded along the azimuthal line.", + "texts": [ + " The onset of convection was also examined by changing the density of E. gracilis. The emergence frequency for convection in a series of experiments depended on the initial conditions, i.e., a difference was seen between the experiments with uniform and localized initial distributions of E. gracilis. This difference suggests that the bioconvection of E. gracilis is bistable. An annular container with an outer radius R \u00bc 25mm and an inner radius r \u00bc 20mm composed of glass was used to achieve the periodic boundary conditions in the azimuthal direction [Fig. 1(a)]. The radial width and the suspension depth d were both fixed to 5mm, which was larger than the critical depth over which the typical wavelength did not significantly depend on depth in the thin cylinder container used by Suematsu et al.13) Using the annular container, we can isolate basic types of localized convection patterns from Journal of the Physical Society of Japan 83, 043001 (2014) http://dx.doi.org/10.7566/JPSJ.83.043001 Letters 043001-1 \u00a92014 The Physical Society of Japan the complex two-dimensional spatiotemporal behavior of localized convection cells described by Suematsu et al", + " gracilis, but the top of the container was covered with glass to ensure that the solvent (distilled water) does not evaporate significantly during the duration of the experiment. All experiments were performed at room temperature (20\u201325 \u00b0C) at night to exclude the effect of the circadian rhythm of E. gracilis. The spatiotemporal pattern of bioconvection was recorded from the top using a digital video camera (JVC GC-PX1). Grayscale images along the center of the annular region [the round line indicated in Fig. 1(b)], l\u00f0 ; t\u00de, were extracted using ImageJ imaging software. The measured values were normalized using the dark region [where l\u00f0 ; t\u00de \u00bc 1] and bright region [where l\u00f0 ; t\u00de \u00bc 0]. Two types of initial states were prepared: one in which the density of E. gracilis was spatially uniform (uniform state) and the other in which the density was spatially localized (localized state). To achieve the localized state, we generated a dark region using a strip of black paper with a width of 5mm, and a suspension with a high E" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001688_j.physe.2010.07.016-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001688_j.physe.2010.07.016-Figure2-1.png", + "caption": "Fig. 2. Schematic picture of a microtubule supported by linear and torsional springs at both ends.", + "texts": [ + " The constants C1\u2013C4 and b can be found from the boundary conditions. To study the microtubule vibration, different boundary conditions such as simply supported ends, clamped ends, etc. can be used. It should be noted that neither of the mentioned boundary conditions represents the actual boundary conditions of the microtubule in a living cell. In this study, we assume general boundary conditions at both ends of the microtubule. We assume that the ends of a microtubule are supported by linear and torsional springs as shown in Fig. 2. The boundary conditions can be stated as EI @2w @x2 \u00fekt1 @w @x \" # x \u00bc 0 \u00bc 0, @ @x EI @2w @x2 ! \u00fek1w \" # x \u00bc 0 \u00bc 0, \u00f032\u00de EI @2w @x2 \u00fekt2 @w @x \" # x \u00bc l \u00bc 0, @ @x EI @2w @x2 ! \u00fek2w \" # x \u00bc l \u00bc 0, where l is the microtubule length. Applying the above boundary conditions to Eq. (31) yields EI\u00f0 C1b\u00feC3b\u00de kt1\u00f0C2\u00feC4\u00de \u00bc 0, \u00f033\u00de EI\u00f0 C2b 3 \u00feC4b 3 \u00de\u00fek1\u00f0C1\u00feC3\u00de \u00bc 0, \u00f034\u00de EI\u00f0 C1bcosbl C2bsinbl\u00feC3bcoshbl\u00feC4bsinhbl\u00de \u00fekt2\u00f0 C1 sinbl\u00feC2 cosbl\u00feC3 sinhbl\u00feC4 coshbl\u00de \u00bc 0, \u00f035\u00de EI\u00f0C1b 3 sinbl C2b 3 cosbl\u00feC3b 3 sinhbl\u00feC4b 3 coshbl\u00de k2\u00f0C1 cosbl\u00feC2 sinbl\u00feC3 coshbl\u00feC4 sinhbl\u00de \u00bc 0: \u00f036\u00de For nontrivial solution of the constants C1\u2013C4, the determinant formed by their coefficients is set equal to zero: EIb kt1 EIb kt1 k1 EIb3 k1 EIb3 EIbcosbl kt2 sinbl EIbsinbl\u00fekt2 sinbl EIbcoshbl\u00fekt2 sinhbl EIbsinhbl\u00fekt2 coshbl EIb3 sinbl k2 cosbl EIb3 cosbl k2 sinbl EIb3 sinhbl k2 coshbl EIb3 coshbl k2 sinhbl \u00bc 0: \u00f037\u00de The expression of damped-frequencies (od \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x=mc\u00feb 4EI=mc\u00de \u00f0a=2mc\u00de 2 q ) for different boundary conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003478_iccas.2013.6704123-Figure12-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003478_iccas.2013.6704123-Figure12-1.png", + "caption": "Fig. 12. Kinematics of wheel loader", + "texts": [ + " The queue lines rated highly (in this case is small) are selected and some queue lines with low rated score are deleted. When there are queue with high score, the several queue lines in the high score ones are inherited as keeping its condition. Repetitive operation in this algorithm ends with in given number of generation. When an operation is done to a queue line and the rated score of the queue is small, the operation is deleted, then the wheel loader with motion by the queue line stops at some position and direction. The kinematics model is shown in Fig.12 and the rep resentative point which indicate the position of the wheel loader is P which is called imaginary point in this paper. Parameters and letters used in Fig.l2 are shown in Table.II. Then, angular velocity (w) of the vehicle is given as w = v jR and the radius of the circle which is the trajectory of the wheel loader is given (2) then w become (3). I R= - tan P. 2 v x tan \ufffd w = ---:--=- I (2) (3) When the wheel loader is the position and posture of Pb the next ones are represented as following (4) Where, 8t is period of sampling time" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002345_j.jbiosc.2012.05.021-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002345_j.jbiosc.2012.05.021-Figure1-1.png", + "caption": "FIG. 1. Schematic diagram of the cell structure.", + "texts": [ + "5 mol/L H2SO4 solution, and then deionized water again. Each stepwas done for 1 h (14). The cathode and themembranewere hot pressed at T\u00bc 135 C and pressure of P\u00bc 5MPa for a time of t \u00bc 3 min. The anode was attached to the membrane without hot pressing. The anode, membrane and cathode were sandwiched between the anode and cathode current collectors. Cell structure The cell prepared in this study was an open air cathode cell of which the electrodes were vertically arranged with an 84 cm3 anode chamber as shown in Fig. 1. The anode and the membrane with the cathode were sandwiched between two stainless steel current collectors that were 1 mm thick with open holes. The cathode current collector was covered with a Gore-Tex sheet to prevent the membrane from drying in the air. Anode medium The yeast extract, YE, used in this experiment was from Nihon Pharmaceutical Co., Ltd. (Tokyo, Japan). A specific amount of YE was dissolved in 70 mL of 50 mM phosphate buffer as the anode medium. In some experiments, D-glucose (Wako Pure Chemical Industries, Osaka, Japan) and methylene blue, MB (Nichido, Inc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003692_978-3-642-23681-5_13-Figure13.6-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003692_978-3-642-23681-5_13-Figure13.6-1.png", + "caption": "Fig. 13.6 Stern-tube test rig (Litwin [29]). 1 Main shaft \u00f8100 mm; 2 tested bearing bush; 3 baring sleeve; 4 covers with sealing; 5 static load leaver; 6 roller bearings; A pressure pickup sensor; B touch less distance sensors; C torqmeter", + "texts": [ + " It has been found from experimental measurements and confirmed using CFD that complex film pressure distributions exist over the loaded staves. These distributions are a result of the interaction of the elastic bearing surface deflections with the film pressures. Peak pressures in the bearings are greatly reduced compared with conventional bearings. Wojciech Litwin [29] has extensively studied the performance of water lubricated bearing through his experimental studies. The experimental setup is as shown in the Fig. 13.6. The diameter of the sea water resistant steel shaft was 100 mm and the length being 700 mm. The shaft was connected with a clutch and torque meter to a electric engine so as to vary the rotating speed. The examined bearing is set in the steel sleeve and processed along with it. The sleeve is closed on both ends with the sealed covers which make it possible for the lubricating and cooling water flow inside it. This enables the bearing to work in different pressures up to the pressure of 0.6 MPa (Fig. 13.7). The load is exerted on the bearing in the static or dynamic way. The parameters which could be varied are as follows: \u2022 Shaft rotation speed from 0 to 11 rev/s. \u2022 The radial load from 1 to 5 kN. \u2022 The pressure of the water supplying the bearing, varying from 0 to 0.6 MPa. These conditions are similar to the real small ship main shaft bearing. The pressure measuring (Fig. 13.6) system was a pressure sensor mounted inside the shaft and the signals were transmitted using a wireless system. The trajectory of the shaft was determined by having two pairs of non-contact transducers with an accuracy of 1 lm on either side of the shaft (Fig. 13.8). The second test rig called the Propeller Shaft bearing test stand is similar to the previous test rig except for the number of bearings to be tested are two. This test rig has the capability of measuring the fluid film pressure in the mid plane of the bearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0000443_acc.2008.4587058-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0000443_acc.2008.4587058-Figure2-1.png", + "caption": "Fig. 2 - Damaged Generic Aircraft", + "texts": [ + " The stability margins are generally functions of \u0393\u03b2 2 0 . The persistent excitation \u03b2 2 0 can be computed from Eq. (27) within a given time window. Using this value, the adaptive gain \u0393 can be calculated and used for adaptation for the next time window. This process is repeated until \u0393 should reach a steady state value when the weights no longer vary. To illustrate the bounded linear stability analysis method, a simulation was performed for a damaged twin-engine generic aircraft with 25% of the left wing missing [2], as shown in Fig. 2. The hybrid adaptive control is implemented in a flight control to track a pitch doublet. Figure 3 is a plot of the pitch rate tracking error. Without adaptation (\u0393 = 0, R = 0), the tracking performance of the flight control is quite poor as the tracking error is large. With the direct MRAC alone (\u0393 = 104, R = 0), the tracking error becomes smaller but high frequency contents also appear. This is consistent with the closed-loop pole analysis. With the hybrid adaptive control (\u0393 = 104 , R = 104I), the tracking error is significantly reduced along with the high frequency contents" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002986_978-3-642-29308-5-Figure2.35-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002986_978-3-642-29308-5-Figure2.35-1.png", + "caption": "Fig. 2.35 The three levels of the Daniels co-axial escapement", + "texts": [ + " It is the masterpiece of Dr. George Daniels (1926\u2013) (Wikipedia 2009). Dr. Daniels is a professional horologist with many achievements. Besides inventing the co-axial escapement, he is also the author of several books on mechanical watch movement (Cecil Cluttoh and George Daniels 1979; George Daniels 1981; George Daniels 2011) and the former president of the Horological Institute. Figure 2.34 shows the model of the Daniels co-axial escapement. It is more complicated than the Swiss lever escapement and has three levels. Figure 2.35 shows the three levels of the escapement. On Level 1, the balance wheel contacts the pallet fork. The guard pins are also on this level. The escape wheel has two levels, one for the inner escape wheel and the other for the outer escape wheel, with 12 teeth on each level. On Levels 2 and 3, the two levels of the pallet fork contact the two levels of the escape wheel. As shown in Fig. 2.36, the Daniels co-axial escapement has six shocks in a cycle. Figure 2.36a shows the first shock. It occurs when the semi-circular impulse-pin on the balance contacts the entry pallet of the pallet fork to unlock the escape wheel" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001213_s1874-1029(08)60091-9-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001213_s1874-1029(08)60091-9-Figure3-1.png", + "caption": "Fig. 3 The state variables xh trajectories", + "texts": [], + "surrounding_texts": [ + "No. 6 WEI Qing-Lai et al.: Data-based Optimal Control for Discrete-time Zero-sum \u00b7 \u00b7 \u00b7 689\nStep 1. Give the boundary condition xh(0, l) = f (l) and xv(k, 0) = g(k). Let P0 = 0, K0 = 0, and L0 = 0. Give the computation accuracy \u03b5.\nStep 2. According to the N sampling points, compute\nZN and Y\u0302 N according to (83) and (84). Step 3. Compute hi according to (82) and Hi according to (84) through the Kronecker method. Step 4. Compute the feedback control laws by\nKi+1 = (Hi uu \u2212 Hi uw(Hi ww)\u22121Hi wu)\u22121\u00d7 (Hi uw(Hi ww)\u22121Hi wx \u2212 Hi ux) (87)\nand\nLi+1 = (Hi ww \u2212 Hi wu(Hi uu)\u22121Hi uw)\u22121\u00d7 (Hi wu(Hi uu)\u22121Hi ux \u2212 Hi wx) (88)\nStep 5. If\n\u2016 hi+1 \u2212 hi \u2016\u2264 \u03b5 (89)\nexit; otherwise, go to Step 6. Step 6. Set i = i + 1, go to Step 2.\n4 Neural network implementation\nIn this subsection, neural networks are constructed to implement the iterative ACD algorithm. There are several ACD structures that can be chosen[27]. As HDP structure is basic and convenient to realize, we will use it to implement the iterative ACD algorithm.\nAssume the number of hidden layer neurons is denoted by l, the weight matrix between the input layer and hidden layer is denoted by V , and the weight matrix between the hidden layer and output layer is denoted by W . Then, the output of three-layer neural network is represented by\nF\u0302 (X, V, W ) = WT\u03c3(V TX ) (90)\nwhere \u03c3(V TX ) \u2208 Rl, [\u03c3(z)]i = ez\ni \u2212 e\u2212zi ez i + e\u2212zi , i = 1, \u00b7 \u00b7 \u00b7 , l, are\nthe activation function. The neural network estimation error can be expressed by\nF (X ) = F (X, V \u2217, W \u2217) + \u03b5(X ) (91)\nwhere V \u2217 and W \u2217 are the ideal weight parameters, and \u03b5(X ) is the reconstruction error.\nHere, there are three neural networks, which are critic network, action network u, and action network w. All the neural networks are chosen as three-layer feedforward networks. The whole structure diagram is shown in Fig. 1. The utility term in the figure denotes xT(k, l)Qx(k, l) + uT(k, l)Ru(k, l) + wT(k, l)Sw(k, l).\n4.1 The critic network\nThe critic network is used to approximate the Hamilton function H(k, l). The output of the critic network is denoted as\nH\u0302i(k, l) = WT ci\u03c3(V T ci x(k, l)) (92)\nThe target function can be written as\nHi+1(k, l) = H+ i (k, l)+\n[ xT(k, l) uT i (k, l) wT i (k, l) ]\u23a1\u23a3Q 0 0 0 R 0 0 0 S \u23a4 \u23a6 \u23a1 \u23a3 x(k, l) ui(k, l) wi(k, l) \u23a4 \u23a6\n(93)\nThen, we define the error function for the critic network as\neci(k, l) = H\u0302i+1(k, l) \u2212 Hi+1(k, l) (94)\nAnd, the objective function to be minimized in the critic network is\nEci(k, l) = 1\n2 e2 ci(k, l) (95)\nSo the gradient-based weight updating rule[39] for the critic network is given by\nwc(i+1)(k, l) = wci(k, l) + \u0394wci(k, l) (96)\n\u0394wci(k, l) = \u03b1c [ \u2212\u2202Eci(k, l)\n\u2202wci(k, l)\n] (97)\n\u2202Eci(k, l) \u2202wci(k, l) = \u2202Eci(k, l)\n\u2202H\u0302i(k, l)\n\u2202H\u0302i(k, l) \u2202wci(k, l) (98)\nwhere \u03b1c > 0 is the learning rate of critic network and wc(k, l) is the weight vector in the critic network.\n4.2 The action networks\nAction networks are used to approximate the iterative optimal controls. There are two action networks, which are used to approximate the optimal controls u and w, respectively.\nFor the action network that approximates the control u(k, l), state x(k, l) is used as the input to create the optimal control and u(k, l) is used as the output of the network. The output can be formulated as\nu\u0302i(k, l) = WT ai\u03c3(V T aix(k, l)) (99)\nSo, we can define the output error of the action network as\neai(k, l) = u\u0302i(k, l) \u2212 ui(k, l) (100)\nwhere ui(k, l) is the target function that can be described by\nui(k, l) = (Hi ww \u2212 Hi wu(Hi uu)\u22121Hi uw)\u22121\u00d7 (Hi wu(Hi uu)\u22121Hi ux \u2212 Hi wx)x(k, l) (101)\nwhere Hi can be obtained according to Kronecker product in (85).\nThe weighs in the action network are updated to minimize the following performance error measure:\nEai(k, l) = 1\n2 e2 ai (102)", + "690 ACTA AUTOMATICA SINICA Vol. 35\nThe weight updating algorithm is similar to the one for the critic network. By the gradient descent rule, we can obtain\nwa(i+1)(k, l) = wai(k, l) + \u0394wai(k, l) (103)\n\u0394wai(k, l) = \u03b2a [ \u2212\u2202Eai(k, l)\n\u2202wai(k, l)\n] (104)\n\u2202Eai(k, l) \u2202wai(k, l) = \u2202Eai(k, l) \u2202eai(k, l) \u2202eai(k, l) \u2202ui(k, l) \u2202ui(k, l) \u2202wai(k, l) (105)\nwhere \u03b2a > 0 is the learning rate of the action network. For the action network w that approximates the control w(k, l), state x(k, l) is used as the input to create the optimal control and w(k, l) is used as the output of the network. The target of w action network can be expressed as\nwi(k, l) = (Hi ww \u2212 Hi wu(Hi uu)\u22121Hi uw)\u22121\u00d7 (Hi wu(Hi uu)\u22121Hi ux \u2212 Hi wx)x(k, l) (106)\nAll the update rules of w action network are the same as the update rules of u network and it is omitted here.\n5 Simulation\nIn this section, the proposed method is applied to an air drying process control. Our example is a modification of Example 1 in [40] and extends the variable space to the infinite horizon.\nThe dynamical processes can be described by the following Darboux equation:\n\u22022x(s, t)\n\u2202s\u2202t = a1\n\u2202x(s, t)\n\u2202t + a2\n\u2202x(s, t)\n\u2202x + a0x(s, t) +\nbu(s, t) + cw(s, t) (107)\nwith the initial and boundary conditions\nxh(0, t) = { 0.5, t \u2264 4 0, t > 4 , xv(s, 0) = { 1, s \u2264 4 0, s > 4 (108)\nwhere x(s, t) is an unknown function, a0, a1, a2, b, and c are real coefficients, u(s, t) and w(s, t) are the input functions. The variable x means the humidity, which is the system state, s means the location of the air, and t is the processing time.\nLet a0 = 0.2, a1 = 0.3, a2 = 0.1, b = 0.3, and c = 0.25. The quadratic performance index function is formulated as\nV =\n\u222b \u221e\nt=0\n\u222b \u221e\ns=0\n{ Qx2(s, t)+Ru2(s, t)+Sw2(s, t) } dsdt\n(109)\nThe discretization method for system (107) is similar to the method in [40]. Suppose that the sampling periods of the digital control system are chosen as X = 0.1 cm and T = 0.1 s. Following the methodology presented in [40], we can compute the discretized system equation (1) as[\nxh((k + 1)X, lT ) xv(kX, (l + 1)T )\n] = [ 0.7408 0.2765 0.0952 0.9048 ] [ xh(kX, lT ) xv(kX, lT ) ] +[\n0.0259 0\n] \u00d7 u(kX, lT ) + [ 0\n0.0564\n] w(kX, lT )\n(110)\nwith the boundary conditions\nxh(0, lT ) = { 0.5, l \u2264 40 0, l > 40 , xv(kX, 0) = { 1, k \u2264 40 0, k > 40\n(111)\nand the discredited performance index function as V = \u221e\u2211\nk=0 \u221e\u2211 l=0 { Qx2(Xk, T l) + Ru2(Xk, T l) + Sw2(Xk, T l) } (112)\nWe implement the iterative algorithm at (k, l) = (0, 0). We choose three-layer neural networks as the critic network, the action network u, the action network w with the structures 2-8-1, 2-8-1, and 2-8-1, respectively. The initial weights of action networks and critic network are all set to be random in [\u22120.5, 0.5]. Then, the critic network and the action network are trained for i = 50 times so that the given accuracy \u03b5 = 10\u22126 is reached. In the training process, the learning rate \u03b2a = \u03b1c = 0.05. The evaluating point number N = 40 for every iteration and choose the small white noise as \u03be1(0, 0.01) and \u03be2(0, 0.01). The convergence curve of the performance index function is shown in Fig. 2. Then, we apply the optimal control to the system for k = 40, l = 40 time steps and obtain the following results. The state trajectories are given as Figs. 3 and 4. The control curves are given as Figs. 5 and 6, respectively.", + "No. 6 WEI Qing-Lai et al.: Data-based Optimal Control for Discrete-time Zero-sum \u00b7 \u00b7 \u00b7 691\nFrom the simulation results, we can see that the proposed iterative ACD algorithm in this paper obtains good effects. In [40], Tsai just studied the model-based optimal control in the finite horizon. In this paper, using the iterative ACD algorithm, the optimal control scheme for 2-D system in the infinite horizon can also be obtained without the system model. So the proposed algorithm in this paper is more effective than the method in [40] for industry process control.\n6 Conclusion\nIn this paper, we proposed an effective iterative algorithm to find the optimal controller of a class of discretetime two-person zero-sum games for Roesser types 2-D systems. The proposed ACD algorithm allows to be implemented without the system model. Stability analysis of the 2-D systems was presented and the convergence property of the performance index function was also proved. The simulation study has successfully demonstrated the upstanding performance of the proposed optimal control scheme for the 2-D systems.\nReferences\n1 Jamshidi M. Large Scale Systems: Modeling, Control, and Fuzzy Logics. Amsterdam: The Netherlands Press, 1982\n2 Chang H S, Marcus S I. Two-person zero-sum Markov games: receding horizon approach. IEEE Transactions on Automatic Control, 2003, 48(11): 1951\u22121961\n3 Chen B S, Tseng C S, Uang H J. Fuzzy differential games for nonlinear stochastic systems: suboptimal approach. IEEE Transactions on Fuzzy Systems, 2002, 10(2): 222\u2212233\n4 Nian Xiao-Hong, Cao Li. Design of optimal observer and optimal feedback controller based on differential game theory. Acta Automatica Sinica, 2006, 32(5): 807\u2212812 (in Chinese)\n5 Nian Xiao-Hong. Suboptimal strategies of linear quadratic closed-loop differential games: a BMI approach. Acta Automatica Sinica, 2005, 31(2): 216\u2212222\n6 Bertsekas D P. Convex Analysis and Optimization. Boston: Athena Scientific, 2003\n7 Goebel R. Convexity in zero-sum differential games. SIAM Journal of Control and Optimization, 2001, 40(5): 1491\u22121504\n8 Altman E, Basar T. Multiuser rate-based flow control. IEEE Transactions on Communications, 1998, 46(7): 940\u2212949\n9 Basar T, Olsder G J. Dynamic Noncooperative Game Theory. New York: Academic Press, 1982\n10 Basar T, Bernhard P. H\u221e Optimal Control and Related Minimax Design Problems. Boston: Birkhauser Press, 1995\n11 Hua X, Mizukami K. Linear-quadratic zero-sum differential games for generalized state space systems. IEEE Transactions on Automatic Control, 1994, 39(1): 143\u2212147\n12 Wei G, Feng G, Wang Z. Robust H\u221e control for discretetime fuzzy systems with infinite-distributed delays. IEEE Transactions on Fuzzy Systems, 2009, 17(1): 224\u2212232\n13 Werbos P J. Approximate dynamic programming for realtime control and neural modeling. Handbook of Intelligent Control: Neural, Fuzzy, and Adaptive Approaches. New York: Van Nostrand Reinhold, 1992\n14 Xu Jian-Ming, Yu Li. H\u221e control for 2-D discrete state delayed systems in the second FM model. Acta Automatica Sinica, 2008, 34(7): 809\u2212813\n15 Uetake Y. Optimal smoothing for noncausal 2-D systems based on a descriptor model. IEEE Transactions on Automatic Control, 1992, 37(11): 1840\u22121845\n16 Owens D H, Amann N, Rogers E, French M. Analysis of linear iterative learning control schemes \u2014 a 2D systems/repetitive processes approach. Multidimensional Systems and Signal Processing, 2000, 11(1-2): 125\u2212177\n17 Sulikowski B, Galkowski K, Rogers E, Owens D H. Output feedback control of discrete linear repetitive processes. Automatica, 2004, 40(12): 2167\u22122173\n18 Li C J, Fadali M S. Optimal control of 2-D systems. IEEE Transactions on Automatic Control, 1991, 36(2): 223\u2212228\n19 Liu D R, Javaherian H, Kovalenko O, Huang T. Adaptive critic learning techniques for engine torque and air-fuel ratio control. IEEE Transactions on Systems, Man, and Cybernetics, Part B: Cybernetics, 2008, 38(4): 988\u2212993\n20 Al-Tamimi A, Abu-Khalaf M, Lewis F L. 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IEEE Transactions on Systems, Man, Cybernetics, Part B: Cybernetics, 2008, 38(4): 937\u2212942" + ] + }, + { + "image_filename": "designv11_3_0002113_09544062jmes2181-Figure3-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002113_09544062jmes2181-Figure3-1.png", + "caption": "Fig. 3 Kamm\u00fcller\u2019s reverse pumping mechanism [13]", + "texts": [ + " There have been a variety of mechanisms proposed to explain this reverse pumping [1]. However, most of these have been discounted on the basis of experiments. A large number of experiments have shown that the reverse pumping is associated with the asperities on the lip surface [8\u201312] and can be explained by a mechanism attributed to Kamm\u00fcller [13] and M\u00fcller [14]. (Actually, Qian [15] proposed the same mechanism 2 years before Kamm\u00fcller and M\u00fcller, but as it was published in a Chinese journal, it was unknown in the West.) This mechanism is illustrated in Fig. 3. As the shaft rotates, shear stresses on the lip surface cause it to deform in the circumferential direction. The asymmetric macro-geometry of the lip (smaller angle on the air side than on the liquid side) causes Proc. IMechE Vol. 224 Part C: J. Mechanical Engineering Science JMES2181 at WAYNE STATE UNIVERSITY on March 21, 2015pic.sagepub.comDownloaded from the deformation to be non-uniform. In a well-designed seal, the maximum displacement of the surface occurs at a location closer to the liquid side than to the air side" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003546_s0263574710000664-Figure1-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003546_s0263574710000664-Figure1-1.png", + "caption": "Fig. 1. An ATRON module (left) and various assembled structures (right).", + "texts": [ + " Then, the main contribution is presented in two parts: first, Section 3 presents our embedded software platform for robust selfreconfiguration followed by Section 4, which presents our high-level language and describes how it compiles to our target platform. The effectiveness of our approach is documented by the experiments in Section 5 and the systematic analysis using simulated modules in Section 6. Last, Section 7 concludes the paper. Our running example throughout the paper is self-reconfiguration of the ATRON robot from a flat \u201c8-like\u201d shape to a car-like shape, both shown in Fig. 1 along with a snake shape. The snake shape is used as a target configuration in specific experiments. We now present related work on self-reconfiguration, programming languages for modular robot ensembles, and reversible programming. The related work is followed by background information on the ATRON robot and details on an initial, motivating experiment with the communication capabilities of the physical ATRON modules. Off-line planning of self-reconfigurable robots has been studied for a large number of different robotic systems", + " Nevertheless, as demonstrated by Yokoyama et al., the issue of reversing state and control has been solved in general, and the same principles could be used in our language. We note that the work by Yokoyama et al. has been highly inspirational for making a reversible language for self-reconfiguration, in particular for details such as designing the operation rotateFromToBy (which by design is reversible) as opposed to simply rotating to a specific degree. The ATRON self-reconfigurable modular robot (Fig. 1) is a 3D lattice-type system.24 Each unit is composed of two hemispheres, which rotate relative to each other, giving the module one degree of freedom. Connection to neighboring modules is performed by using its four actuated male and four passive female connectors, each positioned at 90 degree intervals on each hemisphere. The likewise positioned eight infrared ports are used to communicate among neighboring modules and to sense distance to nearby objects. Two Atmel ATMega128 micro-controllers (one per hemisphere) linked by an RS-485 serial connection control the hardware" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0001663_s12239-009-0049-6-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0001663_s12239-009-0049-6-Figure2-1.png", + "caption": "Figure 2. McPherson strut system (front suspension).", + "texts": [ + " We found that cornering stiffness at 5 km/h is 6% lower than that at high speeds, and aligning moment stiffness is 8% lower than that at high speeds. Since tire properties change considerably at low speeds, tire properties are of consequence in steering returnability. Table 1 lists the measured tire parameters; cornering stiffness in the front and rear tires was 1,354 N/deg and 1,232 N/deg respectively, and the pneumatic trail was 22.1 mm. 3.1. Trajectory of the Kingpin Axis A tire connected to the suspension system through a hub bearing is steered freely around a rotating axis called the kingpin axis. In a McPherson strut system (Figure 2), the kingpin axis is defined as a vector between the strut top mounting point and the lower ball joint point at a lower control arm. On a local tire coordinate system, kingpin axis trajectories migrate on a contact patch plane as well as a wheel center plane according to the tire turning angle, even Table 1. Tire parameters measured at flat track machine. Cornering stiffness (Cf, Cr) 1354, 1232 N/deg Camber thrust coefficient (Kf) 141 N/deg Overturning coefficient (Ko) 40 Nm/deg though the kingpin axis is fixed in a global coordinate system", + " for the inner road wheel (2) for the outer road wheel , where (3) The moment (MV) is transferred to the kingpin moment (MKV) by the projection to the kingpin axis inclination composed of the caster angle (\u03c4) and the kingpin angle (\u03c3). Equation (5) gives the real angle of the kingpin axis (\u03bb), which is the angle between the z-axis and the kingpin axis. Therefore, the restoring moment (MKV) due to vertical force is described as the product of the moment (MV) and the kingpin axis unit vector (eK), as in Equation (4). (4) , where (5) (6) Finally, Equation (7) calculates the steering rack force (Cho and Lee, 2004), since MKV is the cross product of the steering rack force (Fs) and the effective arm vector (reff), as shown in Figure 2. (7) The simulation results (Figure 5) show that at zero steering rack displacement, the total steering rack forces of the inner and outer road wheels cancel one another through left and right wheel symmetry. However, as the steering rack moves laterally, load lever arms change according to caster angle and kingpin inclination. Steering rack forces at both the inner and outer wheel increase in the restoring direction. As a result, we found that the vertical force at steering is beneficial to the restoring moment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0002760_s00542-010-1188-4-Figure2-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0002760_s00542-010-1188-4-Figure2-1.png", + "caption": "Fig. 2 Coordinate system of the coupled journal and thrust bearings", + "texts": [ + " The Reynolds boundary condition was also included in the numerical analysis so as to simulate cavitation. The stiffness and damping coefficients of the proposed method were compared with those of the physical perturbation of the bearing center. The effect of tilting motion on the moment and force coefficients was also investigated. 2 Methods of analyses 2.1 Governing equations This study extended the method of Jang and Lee by including rotational degrees of freedom in the perturbed equations to investigate the moment coefficients and tilting effect (Jang et al. 2006). Figure 2 shows the coordinate system of the coupled journal and thrust bearings. The governing equations for the journal bearing and the thrust bearing were obtained by transforming the Reynolds equation into the hz and rh planes, respectively. o Roh h3 12l op Roh \u00fe o oz h3 12l op oz \u00bc R _h 2 oh oh \u00fe oh ot \u00f01\u00de o ror r h3 12l op or \u00fe o roh h3 12l op roh \u00bc r _h 2 oh roh \u00fe oh ot \u00f02\u00de where R is the radius of the journal, _h is the rotational speed of the shaft, h is the film thickness, p is the pressure, and l is the viscosity coefficient" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv11_3_0003931_j.procir.2014.02.038-Figure4-1.png", + "original_path": "designv11-3/openalex_figure/designv11_3_0003931_j.procir.2014.02.038-Figure4-1.png", + "caption": "Fig. 4. Built cylinders in the starting plate.", + "texts": [ + " Prior to determinate the modulation function, which will decrease the energy near the consolidated powder, it is necessary to know the energy required to melt a point surrounded by molten material (the nominal energy density). This nominal energy density is the maximal energy required to melt an elementary volume surrounded by molten material. Then this value will be modulated in function of the proximity to the consolidated powder. To this aim, 32 cylinders were built directly on the starting plate without overhang surface (Fig. 4). The tested parameters are: beam current (I), beam focus, beam speed (v) and two distances between two trajectories (Table 1). In order to limit the number of built cylinders, a design of experiments approach was used. For each distance between two trajectories a L9 tables of Taguchi designs is used. The energy density (U) of each cylinder is calculated with these parameters but also the beam voltage (V = 60 kV) and the built volume (V0): v IVLU 0V (1) with L the distance traveled by the beam. To determinate if the powder is correctly melted, the Young modulus have been measured on a tensile test bench" + ], + "surrounding_texts": [] + } +] \ No newline at end of file