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+[
+    {
+        "image_filename": "designv11_25_0001022_j.jsv.2008.05.023-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001022_j.jsv.2008.05.023-Figure4-1.png",
+        "caption": "Fig. 4. The local coordinate system of gears.",
+        "texts": [
+            " Here, we use damp coefficient Cg to describe the damp force simply: f d \u00bc Cg _S (10) where _S is velocity of S. The impact force fi and the damping force fd are considered as action in the normal direction at the contact face of the gears. The impact force vector Fi and damping vector Fd can be written as Fi \u00bc f in (11) Fd \u00bc f dn (12) where n is a unit normal vector on the contact face of the gears. It depends on the shape of the gear tooth. Here, the impact force vector Fi and damping force vector Fd between the bevel gear act in the x, y, and z directions of the local coordinate system (Fig. 4). Therefore, the shafts receive the force from the axial, lateral, and torsional directions. The gear element is described by the nonlinear impact force Fi, which depends on the backlash B and the nodal displacements Up and Ug on node of the gears. Since the gear element cannot be ARTICLE IN PRESS Q. Gao et al. / Journal of Sound and Vibration 319 (2009) 463\u2013475468 usually expressed as the element matrix, nodal force vectors, which are equivalent to the impact force vector Fi and damping force vector Fd, are given as element nodal force vectors on the node of the gears"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001685_s11668-009-9233-2-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001685_s11668-009-9233-2-Figure2-1.png",
+        "caption": "Fig. 2 The lorry wheel hub system and the failed tapered roller bearing",
+        "texts": [
+            " The load on each wheel is divided by roller width (25.5 mm), and a value for P has been obtained as 1105 N/mm. Maximum Hertz contact pressure has been calculated as 1975 MPa by using Eq 4. Visual Examination The case under investigation involves contact fatigue failure of tapered roller bearing used in a lorry wheel hub system. The failed bearing was used in a BMC lorry. The vehicle was brought for repair to a local car mechanics service. The general appearance of the parts of the failed bearing is shown in Fig. 2. The first observations showed that contact fatigue failure has not formed all around the inner ring. The damaged surface of the inner ring is approximately half of the whole surface area. Spalling-type Table 1 Basic characteristics of tapered roller bearing Main dimension, mm Load capacity, N Weight, kgd D B Dynamic (C) Static (C0) 55 100 35 137000 196000 1.2 surface contact fatigue failure is clearly visible especially on two different regions (Fig. 3), although pitting is noted in regions remote from the two regions where contact fatigue is apparent"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002213_tdc-la.2010.5762870-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002213_tdc-la.2010.5762870-Figure2-1.png",
+        "caption": "Fig. 2. Contribution of the electromagnetic and electrostatic induction effect [11] to tap-off power",
+        "texts": [
+            "00 \u00a92010 IEEE The phase conductors of the considered transmission line are horizontally arranged each having a quadruple bundle configuration (4 x Rail), as depicted in Fig. 1. The transmission line considered during the study is located in the southern region of Brazil and operates at 525 kV. The shield wire within the segment chosen corresponds to a CAA Dotterel wire. The soil resistivity and tower foot resistance considered are 1000 m and 20 , respectively. The line segment where the electromagnetic induction analysis is carried out was represented by PI parameters (LCC \u2013 Line/Cable constants) with 0.5 km length each (Fig. 2). It should be mentioned one constraint encountered with the shield wire along the line. Due to short-circuit withstanding levels, about 2.48% of the total line length (87.9 km), located at the beginning of the sending-end substation and 3.87% located close to the receiving-end substation (with shorter line spans and worst topography though), has the Dotterel code as shield wire (0.375 \u03a9/km dc). This is because near the substations the short-circuit currents present greater values. The rest of the line (93",
+            " During the simulations, line currents of 50, 100, 200, 300, 400 and 500 A (rms), were considered. The phase currents in this line are not constant but vary along the day, hence the above values considered. In this work, it will mainly be analyzed the electromagnetic induction effect rather than the electrostatic effect. The reason for this preference is because some preliminary tests, as it will be shown later, showed that the contribution of the latter type of induction is smaller when compared to the former effect. Figure 2 [11], shows the two type of induction effects present in a Transmission line. It can be seen that the electromagnetic induction has to do mainly with the current present in the phases (variation of the electromagnetic field around the conductor). Thus, inducing both voltages and currents in another conductor (shield wire), whereas the electrostatic effect is related to the coupling effect (capacitance) between phases and also with the ground. As for the tower configuration showed in Fig. 1, it was of our interest to first determine what was the best shield wire circuit configuration producing the highest tap-off power"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001011_13506501jet437-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001011_13506501jet437-Figure2-1.png",
+        "caption": "Fig. 2 Schematic drawing of tribo-test specimens (SCRP/polyester composite)",
+        "texts": [
+            " About four samples of the SSCF were tested; Ft\u2212 L diagrams were plotted; and the tensile properties were calculated. Tensile and compression measurements of neat polyester matrix and SCRP composites (C-SCRP and U-SCRP) were carried out using a tensile test machine (Universal Material tester, WP300) at a cross-head speed of 1 mm/min and a room temperature of 22\u201325 \u25e6C. The tensile and compression strengths and moduli were thus obtained. The tribo test specimens of the SCRP composite were machined into 11 \u00d7 11 \u00d7 20 mm3 with a rubbing surface of 11 \u00d7 11 mm2 (Fig. 2). Adhesive and abrasive wear tests were performed using block-on-disc and block-on-ring tribo testers (Fig. 3), which were described elsewhere [25, 26]. Dry sliding adhesive wear tests were conducted against a smooth stainless steel disc at ambient conditions of temperature and humidity with different normal loads (20\u201380 N) at a sliding velocity of 2.5 m/s for 1800 s. Abrasive tests were also conducted against abrasive paper (water proof SiC, W400), pasted on a rotating cylinder of JET437 \u00a9 IMechE 2008 Proc. IMechE Vol. 222 Part J: J. Engineering Tribology at University of Bath - The Library on June 25, 2015pij.sagepub.comDownloaded from 60 mm diameter. For U-SCRP specimens, sliding tests were conducted for parallel and anti-parallel orientations (PO and APO), as illustrated schematically in Fig. 2. In both adhesive and abrasive tests, the weight loss in the composite pin was determined after each experiment using 1 mg balance (SETRA EL-4105). SEM (JEOL, JSM 840) and optical (light) microscopic images were used to study fibres and the fractured or worn surfaces of composite specimens. For the SEM, fractured or worn surfaces were gold-coated before starting the scanning process using ion sputtering (model JEOL, JFC-1600). The tensile stress\u2013strain diagrams of four similar samples of the SSCF and polyester resin are shown in Fig",
+            " However, when the abrasive particle attacks the 5 or 10 mm long fibre, it does only over a small portion of the fibre length. Thus, the fibre does not get removed easily but takes more time and repeated wear process before it is removed. In a U-SCRP composite, the fibre mats were oriented normal to the sliding direction and were embedded deeply along their length inside the matrix. Thus, the SCF fibres were exposed to abrasion by hard abrasive particles at their ends, as illustrated schematically in Fig. 2. This could offer good resistance to the removal of the material. Figure 14 shows the wear performance of U-SCRP composites, in which the weight loss of a U-SCRP composite in PO was consistently higher than that obtained for a U-SCRP composite in APO, tested under the same test conditions. This is attributed to the fact that the wear debris of the matrix between the parallel fibre mats were fragmented and removed easily by abrasive particles due to a free path ahead of wear debris on the wear track"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003905_s00542-013-2023-5-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003905_s00542-013-2023-5-Figure2-1.png",
+        "caption": "Fig. 2 Groove configuration in conventional hGJB",
+        "texts": [
+            " Section 2 describes the groove pattern of a conventional hGJB and then introduces the size effect of FDBs. Section 3 describes the numerical program used to analyze the appearance and performance of the proposed multi-step eGJB. Section 4 describes the detailed design and characterization of the proposed multi-step hGJB and then introduces the Taguchi method for the optimal design of the proposed multi-step eGJB. Section 5 discusses the numerical characterization of the proposed multi-step eGJB. Finally, Sect. 6 presents some brief concluding remarks. Figure 2 illustrates the structure of the conventional herringbone-grooved cylindrical journal bearing (hGJB) (liu et al. 2010). During operation, the herringbone-grooved pattern pumps the lubricant in the inward direction of the bearing. This increases the pressure within the hGJB, and therefore improves the bearing performance to prevent contact between the rotating and stationary parts (chao and huang 2005). There are two important design criteria for the hGJBs: (1) optimization of oil leakage to increase the lifespan of the hGJBs (Jung and Jang 2011), and (2) optimization of load capacity to increase the spindle stiffness (Yen and chen 2011)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002648_j.mechmachtheory.2011.11.014-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002648_j.mechmachtheory.2011.11.014-Figure10-1.png",
+        "caption": "Fig. 10. A tetrahedral linkage with the IRA guides.",
+        "texts": [
+            " For tetrahedra, there is always an insphere, so all tetrahedra is a base polyhedron for a homothetic linkage. As an example consider the tetrahedron with vertices (0, 0, 0), (100, 0, 0), (70, 60, 0) and (40, 20, 90). It is easy to locate the inscribing sphere by finding the point equidistant to all four faces. Projecting the sphere center on the faces the homothety centers on the faces are obtained. Dropping the perpendiculars to the edges we obtain the polygonal links. Some configurations of the resulting linkage is given in Fig. 10. Other examples of polyhedra with 3-valent vertices only are prisms, hexahedra, dodecahedra with pentagonal faces and all truncated polyhedra. Of course they can be used provided that there is an inscribed sphere. These example geometries just represent the topology and they can be deformed provided that the faces remain tangent to a sphere; for example the base faces of a prism can be made non-parallel, or the side faces may be deformed such that the vertices are not on a cylinder. Alsowe can include in this classmany polyhedrawith 3-valent vertices which also have vertices with higher valence"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002282_978-90-481-9262-5_35-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002282_978-90-481-9262-5_35-Figure2-1.png",
+        "caption": "Fig. 2 Position workspace of a planar 3R manipulator, relative to the tip point of the last link, assuming that the angle \u03b81 of the first revolute joint is restricted to the [\u2212\u03c0/3,\u03c0/3] range. Points corresponding to singularities are indicated in solid lines, and those relative to boundary and interior barriers are indicated with normal vectors on the forbidden side. Configurations 1, 2, and 3, are an example of a boundary barrier, an interior barrier, and a non-barrier singularity, respectively.",
+        "texts": [
+            " (2) can be classified into two broad categories. They can be non-barrier or barrier singularities, depending on whether there exists a trajectory in the neighborhood of q on C, passing through q, whose projection on U traverses \u03c0u(S) or not, respectively. Points corresponding to barrier singularities, in turn, can be classified as boundary or interior singularities, according to whether they occur over \u2202A or over the interior of A, respectively. An example of each one of these singularity types is depicted in Fig. 2, for the particular case of a planar 3R manipulator. We next provide additional criteria to determine which of these singularity types occurs on a given q0 \u2208 S. Let q = q(v) be a parameterization of C in a neighborhood of q0, with q0 = q(v0). Let n0 be the normal to \u03c0u(S) at u0, which can be computed as indicated in [12]. We can determine whether q0 corresponds to a boundary barrier by examining the sign of \u03c8(v) = n0 T(u(v)\u2212u0), (3) 332 A Complete Method for Workspace Boundary Determination for all local trajectories v = v(t) crossing v0 for some t = t0 whose corresponding path u = u(t) is orthogonal to \u03c0u(S) at u0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002121_tasc.2010.2041545-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002121_tasc.2010.2041545-Figure3-1.png",
+        "caption": "Fig. 3. Magnetic field distribution of the induction motor.",
+        "texts": [
+            " It is shown that eddy current induced in the solid rotor can produce a rotating magnetic field, which will interact with the winding magnetic field and produce the electromagnetic torque. The calculated phase voltage with the proposed method and the results from Workshop Problem 30 are compared in Table I. It is shown that the difference between the analytical results and FEM methods is quite small, below 7%, hence verifying that the modified FEM method has a high accuracy. In order to verify the effectiveness of adaptive time-stepping method, an induction motor driven by a pulse-width-modulated (PWM) voltage inverter is calculated, as shown in Fig. 3. The PWM voltage wave form is shown in Fig. 4 and the corresponding time step size by using the adaptive step size algorithm is shown in Fig. 5. The time stepping error when using a fixed step size of s is shown in Fig. 6. With a fixed step size , the time stepping number is 300. Comparatively, with the proposed adaptive time step size algorithm and if the error tolerance is , the time stepping number is 131. The computing time is reduced by about 56% in this example. The FEM formulation with different depths of the model can be used conveniently to address the practical reality where the axial lengths of the iron cores and PM are different"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000525_s00170-006-0874-y-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000525_s00170-006-0874-y-Figure3-1.png",
+        "caption": "Fig. 3 Tooth profile precision deviation \u03b4y (a), helix deviation \u03b4x and valley spacing on helix fa (b)",
+        "texts": [
+            " The rotary motion of the tool \u03d5 2 is dependent on the rotary motion of the generated gear wheel \u03d5 1 in accordance with the transmission ratio of the technological gear, equation (1). Shift c corresponds to the value of the axial feed per wheel turn. Examples of the flanks of the teeth obtained through hobbing simulation is presented in Fig. 2. The flanks of the toothed wheel are composed of a number or regularly distributed surfaces corresponding to the successive layers of the material being removed by the tool blades. The surfaces are concave, therefore the tooth flanks are scaly in character (Fig. 3). It is evidenced by a curvature analysis in CAD of the tooth flanks (Fig. 4). Their depth along the height of the teeth varies. 3 Experimentation 3.1 Workpiece details The characteristic of the gear subjected to tests were introduced in Table 1. The semi-finished toothed wheels were made by die forging. The material was low-carbon steel for carbonizing of the AMS 6265 sort. The blank was of 30\u201335 HRC. The envelopes of the wheels were thermally toughened up to 35\u201341 HRC. 3.2 Experimental details The toothed rim was conventionally hobbed without parallel feed of the cutter axis"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001225_j.cnsns.2009.02.028-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001225_j.cnsns.2009.02.028-Figure8-1.png",
+        "caption": "Fig. 8. Phase portrait at point b where the fixed point A is a weakly attracting one.",
+        "texts": [
+            " The transformation matrix T is composed by 1 kq1k \u00f0Req2; Imq2; kq1kq1\u00de where q1 and q2 are given by q2 \u00bc 2 ffiffi 6 p Vk 9c ffiffiffiffiffiffi 1 k p \u00fe i Vk\u00f01 2 3k\u00de 3x0c 1 1 2 3 k\u00fe i ffiffi 6 p x0 3 ffiffiffiffiffiffi 1 k p 0 BB@ 1 CCA; q1 \u00bc ffiffi 6 p V\u00f01\u00fe2k\u00de 3c ffiffiffiffiffiffi 1 k p 1 k\u00fe 1 0 BB@ 1 CCA; c \u00bc 1 3 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f03 2k\u00de\u00f01\u00fe 2k\u00de\u00f01 k2\u00de q : Introducing the transformation y \u00bc T 1w, _y \u00bc Jy \u00fe 1 2 g\u00f02\u00de\u00f0y\u00de \u00fe 1 6 g\u00f03\u00de\u00f0y\u00de \u00fe O\u00f0y4\u00de; \u00f029\u00de where the Jordan canonical form J is given by J \u00bc T 1JpT \u00bc 0 x0 0 x0 0 0 0 0 k1 0 B@ 1 CA: In (29), the nonlinear vector functions in transformed coordinates are given by g\u00f02\u00de\u00f0y\u00de \u00bc T 1f\u00f02\u00de\u00f0w\u00dejw\u00bcTy; g\u00f03\u00de\u00f0y\u00de \u00bc T 1f\u00f03\u00de\u00f0w\u00dejw\u00bcTy: Assuming that the center manifold has the quadratic form y3 \u00bc 1 2\u00f0h1y2 1 \u00fe 2h2y1y2 \u00fe h3y2 2\u00de, one can reduce (29) into a twodimensional system up to third order _y1 \u00bc x0y2 \u00fe a20y2 1 \u00fe a11y1y2 \u00fe a02y2 2 \u00fe a30y3 1 \u00fe a21y2 1y2 \u00fe a12y1y2 2 \u00fe a03y3 2; _y2 \u00bc x0y1 \u00fe b20y2 1 \u00fe b11y1y2 \u00fe b02y2 2 \u00fe b30y3 1 \u00fe b21y2 1y2 \u00fe b12y1y2 2 \u00fe b03y3 2: \u00f030\u00de Using the 10 out of these 14 coefficients ajk; bjk, the so-called Poincar\u00e9\u2013Lyapunov constant D can be calculated as [33] D \u00bc 1 8x \u00f0\u00f0a20 \u00fe a02\u00de\u00f0a11 b20 \u00fe b02\u00de \u00fe \u00f0b20 \u00fe b02\u00de\u00f0a02 a20 b11\u00de\u00de \u00fe 1 8 \u00f03a30 \u00fe a12 \u00fe b21 \u00fe 3b03\u00de: \u00f031\u00de Fig. 6 illustrates the variation of D with respect to k. Note that based on the value of parameter V, the Andronov\u2013Hopf bifurcation can be supercritical (D > 0) or subcritical (D < 0). Fig. 7 shows the phase portrait corresponding to point a (k \u00bc 0:500, x \u00bc 0:301) on the stability chart (Fig. 5) and the associated pursuit graph. Fixed point A is exponentially attracting here. Fig. 8 depicts the phase portrait associated with point b (k \u00bc 0:500, x \u00bc 0:408) showing a weakly attracting fixed point A. There is a stable limit cycle born (supercritical Andronov\u2013 Hopf bifurcation) around the fixed point A (point c) when x is increased through its critical value xc (phase portrait and pursuit graph are shown in Fig. 9). The pursuit trajectory in global coordinates has two harmonic components. Fig. 10a shows the \u2018\u2018pursuit graph\u201d of five unicycles (initial conditions are chosen slightly away from the equilibrium formation) when x lies slightly above the curve x2\u00f0k\u00de (point c)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure2.32-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure2.32-1.png",
+        "caption": "Figure 2.32 Nyquist diagram showing gain and phase margins",
+        "texts": [
+            " In doing so it passes through the instability point (\u2212180 degrees, \u22121) at a frequency of one radian per second. This system would oscillate constantly at a frequency of 1 radian per second and by definition is \u2018marginally unstable\u2019. The third locus is the aircraft control system example. The locus passes just inside the instability point indicating that the system is stable but the close proximity to that point suggests that the closed loop behavior may be somewhat oscillatory. From the Nyquist diagram, we can also identify the specific gain and phase margins as was done using the Bode diagram. Figure 2.32 shows the Nyquist diagram for the aircraft control example magnified around the instability point indicating the stability margins. Referring to Figure 2.32, the gain margin in dBs is calculated by taking the reciprocal of the distance \u2018x\u2019 on the graph which is the distance from the origin to the point where the locus crosses the horizontal axis. In this case 1/X is approximately 1/0.2. This means that the gain can be increased by a factor of five before instability occurs. This is equivalent to +14 dB as indicated on the original Bode diagram. The phase margin is defined as how much additional phase lag is needed to cause the locus to pass through the instability point at the frequency where the response vector Alternative Graphical Methods for Response Analysis 59 magnitude is 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002777_cdc.2012.6426745-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002777_cdc.2012.6426745-Figure4-1.png",
+        "caption": "Fig. 4. Fixed-wing UAV Ui",
+        "texts": [
+            " , p\u0304mn }, one simple and direct way to approximate it is to utilize an obstacle circle with a center point defined as p\u0304cp = 1 mn \u2211 j p\u0304j , and a radius of maxj |p\u0304cp \u2212 p\u0304j |. | \u2022 | denotes the Euclidean norm. In the remainder of this paper, it is assumed that the set of all obstacles O considered in the environment is covered and modelled by m isolated obstacle circles, i.e., O \u2282 \u222al=1,...,mCl. In this work, we consider a formation of N nonholonomic fixed-wing UAVs where the notations of each UAV Ui for i \u2208 {1, ..., N} are depicted in Figure 4. Let (xi, yi) denotes the inertial position of UAV Ui, \u03c8i represents the heading angle, Xbi and Ybi denote the UAV body axes, X and Y denote the inertial axes, Vi denotes the airspeed, and \u03c9i = \u03c8\u0307i denotes the turning rate of the UAV. The motion of each UAV satisfies kinematic constraint \u2212x\u0307i sin\u03c8i + y\u0307i cos\u03c8i = 0, and has an operating airspeed envelope of 0 \u2264 Vmin \u2264 Vi \u2264 Vmax and a turning-rate envelope of |\u03c9i| \u2264 \u03c9max. Vmin and Vmax denote the minimum and maximum airspeeds, and \u03c9max denotes the maximum turning rate of the UAV"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001668_tac.1968.1098855-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001668_tac.1968.1098855-Figure1-1.png",
+        "caption": "Fig. 1. Amplidyne principle connection diagTams. (a) ,Dithering signal applied to ordinary control winding. (b) Dithering sgnal apphed to speual control Ninding. (c) Dithering signal applied to special yoke winding.",
+        "texts": [
+            " Continuous magnetization of the given magnetic circuit by means of additional time-variable ampere-turns eliminates ambiguity, or hysteresis, between the input and output of an electromechanical amplifier. In practice, however. additional magnetization may be obtained by utilizing one of the free control windings of the amplifier, as shown in Fig. l(a), but as a result an alternating component is thus superimposed on the control voltage due to transformer effect. This component may be eliminated with the aid of a special c0il[l1~1~1 with a different number of pole pairs from that of the control winding (Fig. 1 (b). Still another method of introducing additional magnetization developed is utilization of a winding mounted on the yoke of the machine (Fig. l(c)). No satisfactory explanation is to be found in the literature as to how the presence of the additional ampere-turns relates to the elirnination of the input-output ambiguity. The present paper attempts to explain the mechanism of this latter effect, using a familiar control theory approach, namely finite signal dithering. ASALYSIS OF EFFECT OF DITHERING SIGNAL OK HYSTERESIS COJIPESSATIOK The magnetic circuit of the amplifier ma"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000658_j.snb.2008.08.033-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000658_j.snb.2008.08.033-Figure3-1.png",
+        "caption": "Fig. 3. Schematic illustration of the electrochemical cell: (1) electrode encl",
+        "texts": [
+            " b) Spanish arrays (Array 2) were obtained from D + T Microelectronica A.I.E, Spain. The array, seen in Fig. 2 consists of gold ultramicroelectrode wires bonded and encapsulated to a printed circuit board, with 256 microdiscs of 5 m diameter and 100 m inter-centre distance. The individual electrodes are inlaid with a 1 m recess and so, in this case, the overall system behaves as a recessed sensor, with RE/CE (Ag) and WE (Au) [16]. The assembling of the electrochemical cell using both sets of icroelectrode arrays is shown in Fig. 3. A thin electrolyte-layer 0.1 ml, 0.1 M TBAP in PC) arrangement was used with a PTFE memrane (Millipore) porous membrane (0.45 m pore size and 50 m embrane thickness) covering the microelectrode. The PTFE memrane is manually stretched and held in position by an \u201cO\u201d-ring. nd prevents the liquid electrolyte in contact with the working lectrode from leaking out of the sensor. The PTFE membrane is lso oleophobic and so this minimises the leakage of the solvent nd keeps the pores open to gas transport"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002441_i2mtc.2012.6229339-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002441_i2mtc.2012.6229339-Figure1-1.png",
+        "caption": "Figure 1. Schematic of a solenoid actuator (flow control purposes).",
+        "texts": [
+            " INTRODUCTION The sensorless principle utilizes the transducer characteristics of the actuator device during excitation, namely there is information encoded about some mechanical quantity in its electrical impedance, e.g. position. Since the actuator is used as a sensor and information is extracted from purely electrical signals, any external transducer and other accessories such as mechanical layout, cabling can be saved, greatly reducing the overall system cost and improving robustness. Regarding solenoids, there has been much research done on sensorless plunger position estimation [1]-[9]. Solenoids are limited travel electromechanical converters most commonly used for flow control purposes (Fig. 1), or as relays and on-off contactors etc. In technical literature several strategies are introduced for estimating the position of the solenoid\u2019s plunger. Rahman et. al. [1] proposed the concept of incremental inductance measured on successive PWM (pulse width modulation) periods but position estimation could not be extended to the full plunger travel. In [2] the current waveform is considered as an amplitude modulated signal, hence position information is extracted from ripple and average current caused by PWM excitation",
+            " Additionally, more computational effort is required. In [7] a flux linkage observer for push-pull solenoids is published to estimate the inductance and position of the device. A method for position reconstruction is presented in [9], which reconstructs magnetic fluxes through the use of auxiliary coils attached to the main windings of the actuator. Under certain working conditions varying external forces might be present on the solenoid\u2019s plunger e.g. formulated from fluid pressure in flow control applications (Fig. 1). In technical literature the effect of an externally applied load and its compensation regarding the sensorless principle of solenoids is still an open issue. Experimental tests are neither presented on the load disturbance rejection of the proposed methods. In [1] external forces are considered to be difficult to predict and model, thus omitted. However, [1], [2], [3] recorded the set of inductance and current ripple values at fixed plunger positions and average currents. From these data load compensation might be possible, although information about its magnitude is lost",
+            " As a result, estimation of the applied load is achieved besides sensorless plunger position estimation. With this knowledge of external force, a solenoid might be used even for force or pressure measurement and force control purposes \u201csensorlessly\u201d, thus extending its possible applications. System performance, robustness and cost effectiveness can also be further improved. The introduced method is applicable under PWM (pulse width modulation) drive conditions. A structural picture of a solenoid, used for flow control purposes, is given in Fig. 1. Without loss of generality, current builds up in the winding due to the terminal voltage (usually from PWM). Through the air gap the magnetic force, (related to coil current) and external load (e.g. fluid pressure) immerge the plunger into the housing, thus altering the outflow orifice. These forces are counteracted by the valve return spring. A solenoid actuator\u2019s major input parameters are voltage, external load and temperature [6], and it consists of three main subsystems [6]. The electrical side converts voltage to current, the electromechanical one generates magnetic force from coil current and the mechanical subsystem transfers force to plunger position"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003927_s10846-013-9930-7-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003927_s10846-013-9930-7-Figure2-1.png",
+        "caption": "Fig. 2 Path following formulation problem",
+        "texts": [
+            " The differential equation describing these dynamics is \u03c8\u0308 = c\u03c4u with \u03c4u representing the yawing moment and c denoting a constant related to the aircraft moment of inertia. The objective of this section is to propose a wind field estimation algorithm based on the effect of the latter on light fixed-wing UAVs motion as described previously. In addition, the effectiveness of the estimation algorithm will be tested for a simple straight line path following application and it will be compared with a wind computation method employed in literature. For illustration purposes of the path following problem, let us consider the airplane in Fig. 2. It has a cross track error relative to the desired path, denoted by d, which is related to the airplane yaw angle, yaw rate and yawing moment. As shown in this figure, the inertial path is aligned with the North axis, i.e. along the x axis of the inertial frame. Note that this assumption is not restrictive since any orientation of the desired trajectory is valid. Hence, the rate of change of the cross track error is according to Eq. 2b. Thus, without loss of generality, the airplane dynamics for trajectory following purpose can be defined as d\u0307 = Va sin \u03c8 + wE (3a) \u03c8\u0307 = r (3b) r\u0307 = c\u03c4\u03c8 (3c) According to these equations, wE is separated from the control input by two integrators which prevents the direct cancelation of the wind effect on the airplane motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002850_0022-4898(65)90022-4-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002850_0022-4898(65)90022-4-Figure6-1.png",
+        "caption": "FIG. 6. Schematic drawing of the bevameter, the device for the determination of shear stress--deformation relationship of the soil.",
+        "texts": [
+            " The first series of experiments was rather a verification of the measuring devices, the second one took place when the fraction measurements were performed and the last one when the scale correlations were checked. The measurements aimed at finding out the depth of sinkage of the plates in the soil while different loads were applied. The plates were forced into the soil with the velocity of 2-3 cm/sec. The introductory measurements proved that changes of velocity did not influence the obtained results. The other bevameter, shown in Fig. 6, was used for measuring soil deformation during shear. The shearing plate was made of an evolved segment of scale model pneumatic tyre which was attached to a steel plate of the size 100 x 33 mm. The measured values were H, W, j, z, and z~. The velocity of shearing plate was 2-3 cm/sec which approximately corresponded to the average velocity of soil shearing by the model tyre, and the depth of sinkage was not greater than 3 era. The results of shearing soil for the unit vertical loads in the limits p=0\"092-0 -85 kg/cm ~ are represented in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure5-1.png",
+        "caption": "Figure 5. Static balancing with two identical rotating masses.",
+        "texts": [
+            " (5a) Solving these equations simultaneously, we find the angles 5(1 and :~2 for all positions of the mechanism, and it is now possible to trace out the trajectories of the counterweight centers. If the solution of equations (4a) and (5a) for any position of the mechanism shows that the sine or cosine of any angle will be greater than + I or less than - 1, this proves that one of the radii p is too small. Examine a particular case of static balancing of a mechanism with two rotating masses, in which the trajectories of these masses coincide (Fig. 5), and the counterweights are identical, i.e. m ~ = m z = m ' . Let us combine the point U,y with the center of the circle O. The tracing out of the trajectories of the centers of the counterweights is accomplished easily in this case. Now determine the positions of the counterweights for the ith position of the mechanism. From point As (defined by the position of U~) draw a straight line through O to the intersection with the circumference and find the position of the center of mass of the counterweights (point B~)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003628_icicip.2012.6391478-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003628_icicip.2012.6391478-Figure2-1.png",
+        "caption": "Fig. 2. Schematic of the triple link inverted pendulum",
+        "texts": [
+            " The key of the feedback module is a Trajectory Graph block, this block is embedded by Sfunction which could compute the value of the trajectory and output this value to system as an external disturbance. This disturbance could break the control system\u2019s equilibrium and lead the ADP controller re-learn to the balance point. As the triple link inverted pendulum balancing task is a popular benchmark in the community [15], we use that benchmark to validate and demonstrate our VR learning and control platform. To do that, we first discuss the detailed design strategy of implementing the system in 3D virtual environment. Fig.2 shows a schematic of the triple link inverted pendulum, with three links of variable length mounted on a cart. Here the length of each link L1, L2, L3 is 0.43m, 0.33m, and 0.13m, respectively. The cart is free to move within the bounds of a one-dimensional track while the links are free to rotate only in the vertical plane of the cart and track [15][16]. The equation governing the system can be described as follows: F (q) d2q dt2 = \u2212G(q, dq dt ) dq dt \u2212H(q) + L(q, u) (2) From the above nonlinear dynamical equation, the state- space model can be described as follows: \u00b7 Q(t) = f(Q(t), u(t)) (3) with f(Q((t), u(t)) = [ 04\u00d74 I4\u00d74 04\u00d74 \u2212F\u22121(Q(t))G(Q(t)) ] Q(t) + [ 04\u00d74 \u2212F\u22121(Q(t))[H(Q(t))\u2212 L(Q(t), u(t))] ] (4) and Q(t) = [ x(t) \u03b81(t) \u03b82(t) \u03b83(t) x\u0307(t) \u03b8\u03071(t) \u03b8\u03072(t) \u03b8\u03073(t) ] T (5) The output of eight state variables are: 1) x(t), position of the cart on the track; 2) \u03b81(t), vertical angle of the first link joint to the cart; 3) \u03b82(t), vertical angle of the second link joint to the first link; 4) \u03b83(t), vertical angle of the third link joint to the second link; 5) x\u0307(t), cart velocity; 6) \u03b8\u03071(t), angular velocity of \u03b81(t); 7) \u03b8\u03072(t) , angular velocity of \u03b82(t); and 8) \u03b8\u03073(t), angular velocity of \u03b83(t)[11]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003807_acc.2012.6314790-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003807_acc.2012.6314790-Figure1-1.png",
+        "caption": "Fig. 1. Schematic diagram of the surface vessel showing the control point (CP) and the center of gravity (CG)",
+        "texts": [
+            " So, the initialization of the control does not fail if the vessel is reasonably far from the desired trajectory at the beginning of the motion, and if it is moved by excessive disturbances too far from the desired trajectory during tracking. The equations of motion of the vessel, expressed in boat\u2019s local frame are m11u\u0307\u2212m22vr + d11u = T cos\u03b1, (1) m22v\u0307 +m11ur + d22v = T sin\u03b1, (2) m33r\u0307 + (m22 \u2212m11)uv + d33r = \u2212(T sin\u03b1)L, (3) where u, v, and r are the surge, sway, and yaw speeds, T is propeller force, \u03b1 is rudder angle and L is distance of CG from the propeller location on the vessel (Fig. 1). Here, the position of the control point, a point on the surface vessel other than its center of gravity, is controlled. The use of the control point helps stabilize the orientation of the underactuated surface vessel without the need for direct orientation feedback. A discussion with more details is presented in [12]. A configuration vector x = [xp, yp, \u03b8] T is defined, where xp and yp indicate the global position components of the vessel\u2019s control point, and \u03b8 is the vessel\u2019s orientation. For simplicity, the control point is chosen on the longitudinal 978-1-4577-1096-4/12/$26"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001283_s11029-010-9111-8-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001283_s11029-010-9111-8-Figure2-1.png",
+        "caption": "Fig. 2. A ring un der the ac tion of a fol lower load.",
+        "texts": [
+            " We should also men tion that the in cre men tal re sult ing strains at an nth step of the it er a tive pro - cess are cal cu lated by the for mula Q H D B R V V Vr r r r n r r r r r r t n1 2 1 2 1 2 1 2 1 22 1D D D[ ] [ ] [{[ ( )]= + + - n] - - -( ) }[ ] [ ]R V Vr r n n1 2 1 1D D , DV 0[ ]0 = , DH 0r r1 2 0[ ] = , n = 1, 2, ... . The it er a tive pro cess is con tin ued un til the ful fill ment of the in equal ity U U U[ ] [ ] [ ]n n n+ - <1 e , (27) where U is the global vec tor of nodal dis place ments, K is the Eu clid ean norm in the space of dis place ments, and e is the cal cu - la tion ac cu racy re quired, which is spec i fied a pri ori. Nu mer i cal Re sults As a first ex am ple, we con sider an iso tro pic ring (Fig. 2) (E = 2.1 \u00d7107 and n = 0.3) and a two-layer com pos ite ring (EL = 2.5 \u00d7107 , ET = 106, GLT = 5 \u00d7105 , GTT = 2 \u00d7105 , and n nLT TT= = 0.25) un der the ac tion of a nonuniformly dis trib uted fol lower pres sure p p e( ) ( cos )j j= -0 1 2 . The geo met ri cal pa ram e ters of both rings are r = 100, h = 1, and b = 20. The sub - scripts L and T cor re spond to the re in force ment and trans verse di rec tions of the com pos ite ma te rial. The thick ness of lay ers of the com pos ite ring are as sumed to be the same, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002683_j.finel.2010.08.001-Figure16-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002683_j.finel.2010.08.001-Figure16-1.png",
+        "caption": "Fig. 16. Deep arch: geometry and loading definitions.",
+        "texts": [
+            "0 104 and mass density r\u00bc 0:1. The frame is modelled using overall 160 elements. Fig. 14 shows the time history of the x3 direction displacements in the tip B and elbow A points. Indications about the shape and duration of the applied load are also indicated. A good agreement was found between the calculated results with those of Mata et al. [21] \u00f0J\u00de. Deformed configurations of the structures at the marked equilibrium points are depicted in Fig. 15. Equilibrium states for the deep circular arch shown in Fig. 16 were computed by the two-dimensional finite element formulation. Several authors, Simo and Vu-Quoc [22], Kouhia and Mikkola [23], Cardona and Huespe [24], have analysed the equilibrium paths for such a structure by using one-dimensional finite element in the geometrically nonlinear regime. ve configurations at marked solution points. A 32 equally spaced element mesh for the whole arch is employed. Here we refer to Young\u2019s modulus E\u00bc6 106 and to Poisson\u2019s ratio n\u00bc 0 while the normal area is squared with edge equal to ffiffiffi 2 p "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001121_j.na.2007.02.015-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001121_j.na.2007.02.015-Figure1-1.png",
+        "caption": "Fig. 1. Inverted pendulum.",
+        "texts": [
+            " By approximating the Poincare\u0301 map to O(\u03b5) directly, a general method of Melnikov type for detecting the existence of asymmetric Type II subharmonic orbits outside the homoclinic cycles is presented. c\u00a9 2007 Elsevier Ltd. All rights reserved. MSC: 34C15; 34G25; 37C29 Keywords: Impact oscillator; Non-smooth system; Melnikov method; Subharmonic bifurcation; Poincare\u0301 map 1. Introduction Consider the following general form of periodically excited nonlinear impact oscillator:{ x\u0308 + g(x) = \u03b5 f (t, x, x\u0307, \u03b5), as |x | < 1, x\u0307 7\u2192 \u2212(1 \u2212 \u03b5\u03c1)x\u0307, as |x | = 1, (1.1) which can be used to model an inverted pendulum impacting on rigid walls under external periodic excitation as depicted in Fig. 1. Here |\u03b5| \u2264 \u03b50 1 for some \u03b50 > 0 and the difference 1 \u2212 \u03b5\u03c1 \u2208 (0, 1] is the coefficient of restitution representing energy loss during impact. Being typical piecewise smooth dynamical systems (called PWS systems for short), impact oscillators have attracted the attention of many researchers. In the 1980\u2019s some special forms of system (1.1) with concrete given functions were discussed by Chow, Holmes, Rand, Shaw et al. (see e.g. in [6,7, I Supported by the MOE Research Foundation for Returned Overseas Chinese Scholars, NSFC (China) 10471101 and SRFDP Grants 20050610003"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001257_j.euromechsol.2008.11.006-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001257_j.euromechsol.2008.11.006-Figure8-1.png",
+        "caption": "Fig. 8. Cross section of air and water filled geomembrane tube.\u23a7\u23aa\u23aa\u23aa\u23aa\u23a8",
+        "texts": [
+            " Table 2 presents a comparison between the results of the present study and those obtained by Cotton (2003) for several different parameters of a geomembrane tube filled with water. It is apparent that as the internal pressure of the tube hint increases, the total height of the tube \u03b7max, as well as the tension, increases and the contact length c between the tube and the foundation decreases. The geomembrane tube which is filled with both air and water is the most complicated case in analyzing the geomembrane tubes. In this case, it is assumed that the water head is H and the air pressure is P\u2217 , the water density is \u03c1 and the mass per unit length of the tube is \u03bb (Fig. 8). To obtain the appropriate relations, we have to combine two formulations presented in previous Subsections 3.1 and 3.2. The non-dimensional governing equations, neglecting the tube weight in bottom part of the tube, can be derived as d\u03be(s) ds \u2212 cos ( \u03c8(s) ) = 0, (34) d\u03b7(s) ds \u2212 sin ( \u03c8(s) ) = 0, (35) dn(s) ds = { \u03bc sin(\u03c8(s)) for \u03b7 h, 0 for \u03b7 < h, (36) n(s) d\u03c8(s) ds = { p\u2217 + \u03bc cos(\u03c8(s)) for \u03b7 h, (h \u2212 \u03b7(s)) + p\u2217 for \u03b7 < h, (37) The above governing equations are non-dimensionalized using the following relations: \u03be = X L , \u03b7 = Y L , s = S L , \u03bc = \u03bb \u03c1L , c = C L , n = N \u03c1gL2 , p\u2217 = P\u2217 \u03c1gL "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure5-1.png",
+        "caption": "Figure 5. Clutch components are added as mass, inertia and stiffness components.",
+        "texts": [],
+        "surrounding_texts": [
+            "The clutch and brake parts within the system are modeled as mass, inertia and stiffness components."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.13-1.png",
+        "caption": "Fig. 3.13. Characterization of the anthropomorphic arm at a shoulder singularity for the admissible solutions of the Jacobian transpose algorithm",
+        "texts": [
+            "4 Consider the anthropomorphic arm; a shoulder singularity occurs whenever a2c2 + a3c23 = 0 (Fig. 3.6). In this configuration, the transpose of the Jacobian in (3.38) is JT P = [ 0 0 0 \u2212c1(a2s2 + a3s23) \u2212s1(a2s2 + a3s23) 0 \u2212a3c1s23 \u2212a3s1s23 a3c23 ] . P , if \u03bdx, \u03bdy and \u03bdz denote the components of vector \u03bd along the axes of the base frame, one has the result \u03bdy \u03bdx = \u2212 1 tan \u03d11 \u03bdz = 0, should be increased to reduce the norm of e as much as possible. implying that the direction of N (JT P ) coincides with the direction orthogonal to the plane of the structure (Fig. 3.13). The Jacobian transpose algorithm gets stuck if, with K diagonal and having all equal elements, the desired position is along the line normal to the plane of the structure at the intersection with the wrist point. On the other hand, the end-effector cannot physically move from the singular configuration along such a line. Instead, if the prescribed path has a non-null component in the plane of the structure at the singularity, algorithm convergence is ensured, since in that case Ke /\u2208 N (JT P )"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.2-1.png",
+        "caption": "Fig. 1.2. Combined side force and brake force characteristics.",
+        "texts": [
+            "1 shows the adopted system of axes (x, y, z) with associated positive directions of velocities and forces and moments. The exception is the vertical force F z acting from road to tyre. For practical reasons, this force is defined to be positive in the upward direction and thus equal to the normal load of the tyre. Also D (not provided with a y subscript) is defined positive with respect to the negative y axis. Note, that the axes system is in accordance with SAE standards (SAE J670e 1976). The sign of the slip angle, however, is chosen opposite with respect to the S AE definition, cf. Appendix 1. In Fig. 1.2 typical pure lateral (x =0) and longitudinal (a - 0 ) slip characteristics have been depicted together with a number of combined slip curves. The camber angle 7 was kept equal to zero. We define pure slip to be the situation when either longitudinal or lateral slip occurs in isolation. The figure indicates that a drop in force arises when the other slip component is added. The resulting situation is designated as combined slip. The decrease in force can be simply explained by realising that the total horizontal frictional force F cannot exceed the maximum value (radius of 'friction circle') which is dictated by the current friction coefficient and normal load",
+            " The resulting change in tyre normal loads causes the cornering stiffnesses and the peak side forces of the front and rear axles to change. Since, as we assume here, the fore and aft position of the centre of gravity is not affected (no relative car body motion), we may expect a change in handling behaviour indicated by a rise or drop of the understeer gradient. In addition, the longitudinal driving or braking forces give rise to a state of combined slip, thereby affecting the side force in a way as shown in Fig. 1.2. For moderate driving or braking forces the influence of these forces on the side force Fy is relatively small and may be neglected for this occasion. This means that, for now, the cornering stiffness may be considered to be dependent on the normal load only. The upper left diagram of Fig.l.3 depicts typical variations of the cornering stiffness with vertical load. The load transfer from the rear axle to the front axle that results from a forward longitudinal force FL acting at the centre of gravity at a height h above the road surface (FL possibly corresponding to the inertial force at braking) becomes: A F - h F (1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.29-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.29-1.png",
+        "caption": "Fig. 1.29. Single track model of car trailer combination.",
+        "texts": [
+            " More specifically, we will study the possible unstable motions that may show up with such a combination. Linear differential equations are sufficient to analyse the stability of the straight ahead motion. We will again employ Lagrange's equations to set up the equations of motion. The original equations (1.25) may be employed because the yaw angle is assumed to remain small. The generalised coordinates Y, ~ and 0 are used to describe the car's lateral position and the yaw angles of car and trailer respectively. The forward spee~ dX/dt (= V = u) is considered to be constant. Figure 1.29 gives a top view of the system with three degrees of freedom. The alternative set of three variables v, r and the articulation angle ~o and the vehicle velocity V (a parameter) which are not connected to the inertial axes system (0, X, Y) has been indicated as well and will be employed later on. The kinetic energy for this system becomes, if we neglect all the terms of the second order of magnitude (products of variables): T 1/'2 m(X 2 - + I ~2) + 89 2} +V2/t~ 2 (1.106) The potential energy remains zero: U = 0 (1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003091_aim.2010.5695776-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003091_aim.2010.5695776-Figure2-1.png",
+        "caption": "Fig. 2. Model of segmented foot. Torsional springs are added on ankle and toe joints. ms is toe mass and mf is the mass of whole foot.",
+        "texts": [
+            " The configuration of the walker is defined by the coordinates of the point mass on hip joint and six angles (swing angles between vertical coordinates and each leg, foot angles between horizontal coordinates and each foot, toe angles between horizontal coordinates and each toe), which can be arranged in a generalized vector q = (xh, yh, \u03b11, \u03b12, \u03b11f , \u03b12f , \u03b11t, \u03b12t) T (see Fig. 1). The positive direction of all the angles are counter-clockwise. The segmented foot structure used in this paper is shown in Fig. 2. The foot mass is distributed at two point masses: one at the center of toe, and the other at the center of the rest part of the foot (ms and mf \u2212ms in Fig. 2). We define foot ratio as the ratio of distance between heel and ankle joint to distance between ankle joint and toe tip, namely a/b in Fig. 2, but not a/c, to make it convenient to compare the proposed model with rigid foot model. Note that there is another definition of foot ratio with implications for the evolution of the foot in biomechanical studies, which is calculated by a/c [19]. It is easy to convert the result of a/b in segmented foot model to a/c if toe length is known. In the following paragraphs, we will focus on the Equation of Motion (EoM) of the bipedal walking dynamics of the proposed model. The model can be defined by the rectangular coordinates x, which can be described by the x-coordinate and y-coordinate of the mass points and the corresponding angles (suppose leg 1 is the stance leg): x = [xh, yh, xc1, yc1, \u03b81, xc2, yc2, \u03b82, xc1f , yc1f , \u03b81f , xc2f , yc2f , \u03b82f , xc1s, yc1s, \u03b81s, xc2s, yc2s, \u03b82s] T (1) The walker can also be described by the generalized coordinates q as mentioned before: q = [xh, yh, \u03b11, \u03b12, \u03b11f , \u03b12f , \u03b11t, \u03b12t] T (2) We defined matrix T as follows: T = dq dx (3) Thus T transfers the independent generalized coordinates q\u0307 into the velocities of the rectangular coordinates x\u0307"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000305_t-pas.1975.31993-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000305_t-pas.1975.31993-Figure3-1.png",
+        "caption": "Fig. 3 - PM Motor with 4-pale Stator, 5-tooth Rotor and 4 Phase",
+        "texts": [
+            " BACK EMF COEFFICIENTS We define back EMF coefficients as Kg, represents a self-induced term in the i\" coil, Kgij represents a coupling term, and Kgimrepresents coupling due to the permanent magnet, With the approximation of we find from (lo), (11) and (13a) Kgii - N2 Pli I (1 - 2Kp Po) sin ti - Kp Psin 2 f i ] Kgif = Kp N Po P l j [(sin 5, +sin tj ) + - sin 2 P PO which give Kgi. = Id PIi. [ (1 - 2Kp Po ) sin 5 , - K, P sin 2Si) All the inductance terms and back EMF coefficients are given in terms of permeances. The air gap permeances are solely dependent on the tooth geometry [ 17 1 . Consequently, L's and KgOr can be calculated Once Ps are known. In the following sections we use the foregoing f o m - lation to investigate specific types of motors. 5. SINGLE -STACK PM MOTOR For clarity let us use a 4-pole stator, 5-tooth rotor, 4- phase PM step motor as shown i n Fig. 3. This motor i s chosen because i t i s the simplest 4-phase device whose analytical result can be easily extended to many-toothed devices with minimum effort. Here K = 4 N, = 5 8 e = 58, 1 1 K, --e - 4P0 + P, 4P0 - - where Lo = ;f N Po 3 2 Back EMF Coefficients are from (14) Kgi = - II Llsin f i - L2 sin 2 f i 1 Kglj = I j [ T1(sin t i +s in f j ) + Lzsin (ti + f j )1 L where In order to evaluate the toque, let us suppose for single-phase energization Fi = - F , = -NI . The solution of (4) gives where T, represents a cogging toque and AT includes higher-order terms",
+            " However, based on the author\u2019s dominant permeance components are d-c and first harmonic terms. Therefore, the paper neglects the higher order components in discussing specific cases. The 4th harmonic component was intentionally included to show that the cogging torque for 4-phase PM motors, which are widely used in industry, is due to this component. For other phases, the cogging torque will be due to other component as the discussor pointed out. By single phase energization, the author means that only one winding is energized. In practical motors, Coil 1 and 3 in Fig. 3 are connected in series into a one continuous winding to an appropriate power supply. Ka = 0 in Fig. 6 approximately holds when the air gap flux due to the permanent magnet dominates over that due to the energization flux. Therefore, in practical operations where the energization flux is high a pure sinusoidal torque is not attainable. The idealized mathematical model. developed in the paper gives basic insights into the operation of step motors. Next areas of investigation should include effects of eddy current and non-linear magnetic circuit"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003149_jsen.2010.2041774-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003149_jsen.2010.2041774-Figure2-1.png",
+        "caption": "Fig. 2. A CNT-based alcohol sensor combining two sensing elements.",
+        "texts": [
+            " The resistance of sensing element will recover to the value before the measurement. The recovery time from heating start to full recovery of resistance depends on the CNT sensing element and the heating temperature. The recovery time of CNT sensing element delays the response of alcohol sensor, CNT-based alcohol sensor needs decreasing its response delay due to the recovery of CNT sensing element. A feasible solution is integrating multiple sensing elements with a microprocessor, a time module, and a heating module, and then producing one alcohol sensor (Fig. 2). The time module is used to afford the timing signal to the microprocessor that: 1) connects the signal terminals with one sensing element; 2) control the heating module to heat the other sensing elements; and 3) switch the working state of each sensing element at appropriate time. This kind of alcohol sensor can use one sensing element to measure alcohol concentration, while it simultaneously recovers the other sensing elements (Fig. 2). The response of whole alcohol sensor is speeded up though the recovery time of single CNT sensing element is not deceased. Similarly, integrating more CNT sensing elements into one sensor can increase the measurement frequency that the sensor can achieve. At the same time, this approach raises the complexity of CNT-based alcohol sensors. The above CNT-based alcohol sensor reduces the response delay because of the recovery of CNT sensing element. Its functional component is still the CNT sensing element"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002986_s00170-012-4659-1-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002986_s00170-012-4659-1-Figure12-1.png",
+        "caption": "Fig. 12 Geometry of the machining tool",
+        "texts": [
+            "1 Finishing process for driving tines As shown in Fig. 1, the central line of the driving tine of the SCMW samples is a cylindrical helix. And the equation of the central line is expressed in Eq. 3 [1]. The section of the driving tine is circle. Suppose the diameter is d: x1 \u00bc m cos t y1 \u00bc m sin t z1 \u00bc nt \u00fe np p t p 2 8< : \u00f03\u00de wherem is the radius of the spiral curve and 2n\u03c0 is the pitch of the spiral curve. The schematic diagram of the ECB finishing system designed for the driving tines is shown in Fig. 11. As shown in Fig. 12, the machining tool has six machining areas which match the driving tines. The section of the machining area is circle. Suppose the diameter is D, and the relationship between D and d is expressed as Eq. 4: D \u00bc d \u00fe 2 h \u00f04\u00de where h is the thickness of the insulating cloth. As shown in Fig. 13, the driving tines could be finished by the aid of a four-axis positioning device. The device consists of an XY stage for positioning a machining tool, a rotating mechanism, and a traversing mechanism for positioning a driving wheel"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003334_1.5062408-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003334_1.5062408-Figure1-1.png",
+        "caption": "Figure 1 The structure of the coaxial laser wire welding head [7]",
+        "texts": [
+            " This feature is important in many processes, e.g. laser cladding, laser brazing and laser welding with additive material. Coaxial wire cladding was an interesting opportunity to research the new cladding method. The objective was to improve the accuracy of the wire cladding process and achieve high quality corrosion resistance Inconel 625 coatings by using the new coaxial wire cladding method. The cladding tests were done with a Rofin 4.4 kW diode-pumped Nd:YAG laser and coaxial wire welding head. Figure 1 shows the structure of the coaxial laser head. The mirror optics of the laser head divides the laser beam into three parts, which enables the space for the centric wire feeding lead. The laser beams reflect from the laser head symmetrically at about an 60 degree angle and focus together at the focal length 140.8 mm. The size of the focal point is adjustable with the collimation lens between 0.9 - 1.6 mm. Different beam shapes and sizes in relation to the focal position are shown in Figure 2. The laser spot diameter is about 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002727_ipec.2010.5543188-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002727_ipec.2010.5543188-Figure1-1.png",
+        "caption": "Fig. 1. Definition of vector diagram of IPMSM.",
+        "texts": [
+            " For the sensorless control of IPMSM, the signal injection technique is utilized for the starting algorithm, and the back-EMF based sensorless control is utilized for high speed operation. The proposed smooth transition method utilizes a PLL type position and speed estimator which is used in both the signal injection and the back-EMF based sensorless. The proposed method is is simple and easy to implement. Several experimental drive tests are demonstrated in various load conditions and the experimental results show the effectiveness of the proposed method. Fig. 1 shows the vector diagram of IPMSM, which indicates the relation of two reference frames used in this paper. The voltage equation of an IPMSM in the aligned synchronous reference frame is given as follows : \ufffd ve d ve q \ufffd = \ufffd rs + pLd \u2212\u03c9Lq \u03c9Ld rs + pLq \ufffd \ufffd ied ieq \ufffd + \ufffd 0 \u03c9\u03c8m \ufffd , (1) where ve d, ve q are the d-q axes applied voltages, ied, ieq are d-q axes currents, \u03c9 is the angular speed, rs is the stator winding resistance, Ld, Lq are d-q axes inductances, \u03c8m is the EMF constant, and p is the differential operator. The above voltage equation is usually utilized for a highperformance vector control for PMSM drive with a position sensor. However, in the sensorless control system, (1) cannot be used because the actual rotor position, \u03b8 is unknown. Therefore, the sensorless control is performed in the estimated 978-1-4244-5393-1/10/$26.00 \u00a92010 IEEE frame corresponds to the estimated position, \u03b8\u0302 (misaligned d-q axes in Fig. 1). Let \u0394\u03b8 = \u03b8\u0302 \u2212 \u03b8 and eJ\u0394\u03b8 = \ufffd cos\u0394\u03b8 sin\u0394\u03b8 \u2212 sin\u0394\u03b8 cos\u0394\u03b8 \ufffd . (2) Then, if the estimated angular speed error is significantly small, \u03c9\u0302 \u2212 \u03c9 \u2248 0, the voltage equation of an IPMSM in the misaligned coordinate is written as v\u0302e dq = \ufffd rs \u2212 \u03c9\u0302L\u03b3 + pL\u03b1 \u2212\u03c9\u0302L\u03b2 + pL\u03b3 \u03c9\u0302L\u03b1 + pL\u03b3 rs + \u03c9\u0302L\u03b3 + pL\u03b2 \ufffd i\u0302 e dq +\u03c9\u03c8m \ufffd sin \u0394\u03b8 cos\u0394\u03b8 \ufffd , (3) where L\u03b1 = Ld + Lq 2 \u2212 Lq \u2212 Ld 2 cos 2\u0394\u03b8 L\u03b2 = Ld + Lq 2 + Lq \u2212 Ld 2 cos 2\u0394\u03b8 L\u03b3 = Lq \u2212 Ld 2 sin 2\u0394\u03b8 v\u0302e dq = eJ\u0394\u03b8ve dq, i\u0302 e dq = eJ\u0394\u03b8iedq. For the starting methods from standstill to a certain low speed, various methods are utilized [13]-[19]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000413_0-387-23335-0_3-Figure3.5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000413_0-387-23335-0_3-Figure3.5-1.png",
+        "caption": "Figure 3.5. Schematic illustration of the variation in intensity of a modulated excitation light source (broken line) operating at a frequency = 100 MHz and the resulting phase-shifted luminescence (solid line) generated by a lumophore, which absorbs some of the excitation light, with a lifetime of 10 ns. Under such conditions, from eqn. (13), (|) \" 86.96\u00b0 (s 2.25 ns). \"A\" and \"B\" and \"a\" and \"b\" in this diagram are the measured \"amplitude\" and \"background\" light intensity levels for the lumophore and excitation source, respectively, fi'om which a value for m can be calculated using eqn. (15).",
+        "texts": [
+            " The experimental observables are the shift in phase angle of the emission, (j), and its modulation, m, in intensity, both relative to the phase and modulation of the source. An average decay time, T, can be calculated from the phase angle shift or modulation through either one of the following equations^ : tan (|) = COT (13) and w = (1 + co^^)\"'\" (14) where co is the angular modulation frequency (= 1-n.f, where/= frequency in Hz of the modulated excitation light). A simple illustration of the output of a phase-modulated system is illustrated in figure 3.5. The broken line depicts the sinusoidal variation in excitation light, in this case set at a typical value/= 100 MHz, as a function of time. The solid line shows how the emitted light intensity generated by a fluorophore, with a lifetime, T, (in this case x is set at 10ns), comparable to IZ/'would vary as a function of time if it absorbed some of the excitation light. Due to the finite lifetime of the excited state, the emission will be delayed in time relative to the excitation, which is measured as a phase shift, i), which can be used to calculate x via eqn. (13). The relative amplitude of the emission is also reduced compared to that of the excitation. This process of demodulation allows the parameter, m, to be calculated via the following equation; OT = (A/B)/(a/b) (15) using the data in Figure 3.5. For any value of m a value of T can be calculated via eqn. (14). It should be briefly noted that the above equations hold only if the luminescence decay is described by a single-exponential decay. Interestingly, for FRET-based carbon dioxide sensors, and, for that matter, most FRET-based sensors, e.g. for oxygen, invariably the assumption of mono-exponential excited state decay kinetics is not valid, especially for luminescent dyes dispersed in a heterogeneous medium such as a polymer'\"''\". Fortunately, as we shall see, the variation in the apparent average lifetime, as measured using phase modulation spectroscopy and eqn"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000989_0005-2744(72)90265-3-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000989_0005-2744(72)90265-3-Figure3-1.png",
+        "caption": "Fig. 3. Inhibi t ion of yeast glyceraldehyde-3-phosphate dehydrogenase act ivi ty by ANS with respect to vary ing NAD + concentrat ions, Reaction mixtures contained o.33 mM glyceraldehyde 3-phosphate, 1.28. IO -8 M enzyme. The concentrat ion of NAD + varied from o.o16 to o. 198 raM. Line i, no ANS; Line 2, 0.o83 mM ANS.",
+        "texts": [
+            " Presence of some cationic site in the vicinity of the essential cysteine residue in the active center of glyceraldehyde-3-phosphate dehydrogenase was suggested by several authors14,1'~. This site is, however, inaccessible for ANS, since the dye does not interfere with the interaction of the substrate and the enzyme. The hydrophilic character of the amino acid residues neighbouring the essential cysteine in the pr imary structure of glyceraldehyde-3-phosphate dehydrogenase le may part ly account for a restricted access of the dye to this region of the active center. On the other hand, as seen from Fig. 3, ANS is strictly competitive with NAD \u00f7, the inhibitor constant Ks being o.o55 raM. This value has been found to be in good agreement with the fluorescence ti tration data on the dissociation constant of the enzyme-ANS complex 9. Binding of such a complex molecule as NAD + to a protein is expected to involve contributory interactions arising from different portions of the coenzyme. The importance of the adenine moiety of NAD\u00f7 molecule for the binding to yeast glyceraldehyde-3-phosphate dehydrogenase was established by demonstrating the strictly coenzyme-competitive character of adenine derivatives as inhibitors ~& Purely coenzyme-competitive nature of the binding of both ANS and adenine derivatives suggests the binding of these compounds at the NAD + binding site of the enzyme"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001358_jab.25.2.184-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001358_jab.25.2.184-Figure4-1.png",
+        "caption": "Figure 4 \u2014 A schematic illustrating the variables used to describe ball and stick kinematics at the time the ball was released: ball velocity vector (b), stick tip velocity vector (t), the angle between the stick shaft and horizon (), angle between t and the stick shaft angle (), and the angle between b and t ().",
+        "texts": [
+            " The instantaneous center of rotation was defined as the point of intersection between the helical axis of motion and the plane best-fit to the stick\u2019s motion just before and just after the release of the ball. Tip velocity (t) was calculated given the virtual location of the tip marker and the six degree-of-freedom kinematics of the stick. The linear speed at the tip of the stick was the magnitude of t. The angle of the stick shaft with the horizon (), the angle between the tip velocity and stick shaft (), and the angle between the ball velocity and stick tip velocity () were also calculated at release (Figure 4). A paired Student t test was used to determine if the differences in linear speed between the ball and the stick Figure 2 \u2014 Front and side view photographs of the men\u2019s lacrosse stick heads (AM and BM) and women\u2019s lacrosse stick heads (CF and DF) used in this study. D ow nl oa de d by N ew Y or k U ni ve rs ity o n 09 /1 7/ 16 , V ol um e 25 , A rt ic le N um be r 2 but only by a mean difference of 0.9 m/s (2.1 mph) and 0.4 m/s (0.9 mph), respectively. Although it was not the specific goal of this study, we were able to detect statistically significant differences between the two men\u2019s sticks and the two women\u2019s sticks"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000882_4243_2008_022-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000882_4243_2008_022-Figure3-1.png",
+        "caption": "Fig. 3 Two typical designs of fiber optic sensors. In the upper one, the solid support containing an indicator probe or a chromogenic enzyme substrate forms the cladding of the fiber an is interrogated by an evanescent wave. In the lower one, the material is fixed at the distal end of the fiber. Revised from [18]",
+        "texts": [
+            "1 These sensors rely on the continuous measurement of the intrinsic luminescent properties of a sample, a lot of which were among the first fluorometric sensors reported. Selectivity is achieved by proper choice of analytical wavelengths, time-resolved data acquisition, or by measurement of decay times. Often a fraction of the sample to be sensed is passed through an external loop where luminescence can be recorded. Such sensors are often combined with fiber-optical technology. The fiber acts as a light guide and allows remote spectrometric analysis of any analyte displaying intrinsic luminescence that can be discerned from the background (Fig. 3). Such fiber sensors are referred to as bare-ended fiber sensors, plain fiber sensors, or passive opt(r)odes. Typical examples include the sensors listed in Table 2. Plain sensors display both advantages and disadvantages with respect to quality assurance. In principle, it is usually preferable to use a signal which is generated directly by the analyte, thus avoiding any errors introduced by indirect detection. This is a convenient method, if the fluorescence of the analyte is in the red or near-IR, where the fluorescence background is normally very low, however, the luminescence of most analytes usually has to be excited in the deep UV (< 300 nm), where there is typ- ically a lot of interference, particularly in dense media, such as biological systems",
+            " This situation gave rise to the development of fluorosensors, in which the analytical information is mediated by some sort of indicator chemistry, usually deposited in the form of a thin sensor film. The film is composed of an analyte-permeable polymer that contains the chemically responsive probe. This film can be used in various ways but mostly in the form of a sensor spot as shown in Fig. 2 which is a schematic of the sensing unit of a widely used medical system. Such films may, however, also be deposited inside a reaction bottle, a microwell plate, or at the core of a fiberoptic waveguide as shown in Fig. 3 where a fluorogenic enzyme substrate in a polymeric solid support is placed on either the core or the distal end of the fiber. The support also may contain indicator probes for analytes such as pH, oxygen, and the like. Other examples of such \u201copt(r)odes\u201d are listed in Table 3. A subgroup of such sensors is called \u201creservoir sensors\u201d: In this sensor type a reagent is continuously added to the sample at the tip of an optical fiber. A fluorogenic reaction occurs in the immediate vicinity of the fiber leading to the generated fluorescence being transmitted into it"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000630_s1560354708040060-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000630_s1560354708040060-Figure2-1.png",
+        "caption": "Figure 2",
+        "texts": [
+            " As mentioned, the first step in explaining the motion of TT is to reconcile its motion with the known principles of mechanics. We recall arguments of Del Campo [6] who observed that rising of CM increases the potential energy by about mg\u0303R(1 + \u03b1) \u2212 mg\u0303R(1 \u2212 \u03b1). It can only take place at expense of the kinetic energy 1 2L\u03c9 \u2248 1 2I\u22121 3 (Lz\u0302)2 that is initially accumulated in the z\u0302 component of the angular momentum. Reduction of Lz\u0302 requires action of z\u0302 component of the torque of external forces K. For definiteness we assume, as in Fig. 2 that L is vertically almost parallel to z\u0302. The gravitational force \u2212mg\u0303z\u0302 has zero arm of force. By decomposing the friction force Ff = \u2212\u03bcgnvA = F\u22a5 + F|| into two components one F|| in the plane of z\u0302, 3\u0302, and F\u22a5 orthogonal to this plane we see that the total torque about CM is K = a \u00d7 (FR + F|| + F\u22a5). The torque a \u00d7 (FR + F||) is parallel to the plane of support and only the a \u00d7 F\u22a5, which is in the plane of z\u0302, 3\u0302, has a component in z\u0302 direction. We see also in Fig. 2 that Kz\u0302 is acting in the right direction \u2014 it is reducing the magnitude of L. We can also see from the picture that the torque Kz\u0302 is proportional to sin \u03b8 so it reduces Lz\u0302 most effectively when the inclination of the axis is \u03c0 2 while Kz\u0302 is small in the initial and the final phases of inversion when \u03b8 is close to 0 or \u03c0. REGULAR AND CHAOTIC DYNAMICS Vol. 13 No. 4 2008 These arguments cannot substitute precise mathematical statements about features of TT motion but they help to reconcile them with our everyday experience and serve as a guide how to analyze equations (1)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure4-1.png",
+        "caption": "Figure 4. Static balancing of a mechanism with rotating masses.",
+        "texts": [
+            " The trajectories of the centers of the counterweights correspond to the relationship x' t = a and _z=-' b. However, it is possible to select as trajectories any straight lines (with the exception of parallel lines) intersecting at any angle. Such an arrangement is shown in Fig. 3, where the trajectories are given by the relationships and z i =(x i - a)tg .x - ' = b \";2 The magnitudes a, b, and 5( must be arbitrarily selected. However, from a design point of view, it is more expedient to have a counterweight whose center of mass moves in a circle and not in a straight line. Figure 4 shows the trajectory of the mass center of the moving links and two circular trajectories of the counterweights. Let us write the equations for these trajectories: (St - x l ) 2 + (z't) 2 =p~; ( x . _ S , ) 2 , 2 . + ( z 2 ) = p ~ and assign the magnitudes of Sa, $2, p~ and P2. trigonometric functions of the angles measured between the radius-vectors and the x-axis, in place of the point coordinates. Then the equations of type(4) and (5) assume the forms: lrl~Xn + m ' l ( S 1 - P l c o s ~ 1) - m2(S2 + P 2 C O S 5 ( 2 ) = O ; (4a) i t \u2022 m"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002303_978-94-007-2069-5_58-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002303_978-94-007-2069-5_58-Figure2-1.png",
+        "caption": "Fig. 2 Arrangement of piezoactuators and dependence of their force on displacement",
+        "texts": [
+            " If a feedback between the shaft position and bushing position with respect to the bearing housing is introduced then the critical speed is enlarged to the value MAX D CRIT p KP C 1 : (1) where KP is the gain of the open control loop. The control system does not stabilize the behavior of the journal bearing directly by changing the position of the bearing bushing, but indirectly by changing force that acts on this bushing. Except of the controller gain, the displacement of the bushing depends on stiffness of its connection with the bearing body through rubber seal rings as it is shown in Fig. 2. The dependence of the piezoactuator travel and force on electrical voltage and the dependence of clamping force on bushing displacement are depicted on the right side of Fig. 2. The working point of the electromechanical system results from the voltage which is supplied to the piezoactuator. The gain KP of the open control loop results not only from the setting up of the controller, but from the property of the bushing clamping as well. Properties of the piezoactuator of the P-844.60 type (catalogue values) and measured stiffness of clamping (5.5 106 N/m) gives the bushing travel range which is equivalent to the control variable range. The range of the rotor stable rotational speed is limited by the travel range of piezoactuators and measurement errors of the proximity probes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001954_13506501jet521-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001954_13506501jet521-Figure4-1.png",
+        "caption": "Fig. 4 Position of a temperature measurement and temperature measuring points",
+        "texts": [
+            "1287 mgKOH/g Liquidoid pocking test (synthetic sea water) Rustless Mechanical admixture Naught Water Naught Demulsification number (40-37-3) 54 \u25e6C, min 9\u203230\u2032\u2032 Copper sheet test (100 \u25e6C, 3 h) level 1 emulsions used in the experiment was obtained by mixing the lubricant oil and synthetic sea water by mass percentage. Figures 4(a) and (b) show the position of the temperature measurement and temperature measuring points, respectively. The type of the temperature sensor used was a thermoelectric couple due to its small diameter (less than 1 mm). Such a small size allowed it to be built and mounted in the bearing. Four blind holes were machined in the bearing bush next to the Babbitt alloy (Fig. 4(a)). Thermocouples were mounted in the holes. The sensor was embedded into the housing 1.5 mm below the bearing surface, therefore avoiding direct contact with the Babbitt alloy. The temperature was measured at six different points, four on the bearing (Fig. 4(b)) and the other two on the oil inlet and outlet, respectively. Point 1 was on the vertical line. Point 2 was the location close to the thinnest point in the oil film, with an angle of 30\u25e6 away from the vertical direction. Points 3 and 4 were distributed symmetrically Proc. IMechE Vol. 223 Part J: J. Engineering Tribology JET521 \u00a9 IMechE 2009 at PENNSYLVANIA STATE UNIV on May 23, 2015pij.sagepub.comDownloaded from with an angle of 45\u25e6 to the vertical line. The temperatures were obtained under the steady-state condition to take these measurements, which were deemed to be close to the oil film temperature"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000519_s11071-006-9176-z-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000519_s11071-006-9176-z-Figure3-1.png",
+        "caption": "Fig. 3 Planar double pendulum with impact",
+        "texts": [
+            " (35) The energetic coefficient of restitution e\u2217 [3] can be written as e2 \u2217 = \u2212Wn(ps) \u2212 Wn(pc) Wn(pc) = x \u2212 \u03bck y \u2212 s x + \u03bck y + s ( ps pc \u22121 )2 . (36) Springer From Equation (36), one can compute the separation impulse ps as ps = pc ( 1 + e\u2217 \u221a x + \u03bck y + s x \u2212 \u03bck y \u2212 s ) . (37) The ratio of the angular velocities of separation \u03c9(ps) and approach \u03c9(0) becomes \u03c9s \u03c9a = \u03c9(ps) \u03c9(0) = e\u2217 \u221a x \u2212 \u03bck y \u2212 s x + \u03bck y + s . (38) For no rolling friction moment (s = 0), the ratio \u03c9s/\u03c9a becomes \u03c9s \u03c9a = e\u2217 \u221a 1 \u2212 \u03bck cot \u03b8 1 + \u03bck cot \u03b8 . (39) 3.2 Double pendulum and kinematic coefficient of restitution In Fig. 3, two uniform rigid rods 1 and 2 with lengths L1 and L2 and masses m1 and m2 are joined at point B by a frictionless pin joint in order to form a planar double pendulum. The end of rod 1 pivots around a frictionless pin joint at O . The free end of rod 2 strikes a rough horizontal surface S at point C . The rods have angles of inclination from axis Oy denoted by \u03b81 and \u03b82 and angular speeds of magnitudes \u03c91 = \u03b8\u03071 and \u03c92 = \u03b8\u03072, respectively. The fixed cartesian reference frame x Oyz is chosen. At the impact point C , the coefficients of kinetic and static friction are \u03bck and \u03bcs , the coefficient of rolling friction is s, and the kinematic coefficient of restitution is e",
+            " A relation can also be established between the contact radius r of the pendulum and the energy dissi- pated at impact (Fig. 6). Less energy is dissipated by friction for smaller values of r . For example, for L = 0.200 m, \u03bck = 0.5 = constant, and \u03b8 = 1.044 rad, the ratio (\u03c9s/\u03c9a)/e\u2217 = 0.729 corresponds to r = 0.001 m Fig. 8 Energy variation T function of the coefficient of friction \u03bck for different values of the coefficient s Springer and the ratio (\u03c9s/\u03c9a)/e\u2217 = 0.706 corresponds to r = 0.004 m. 4.2 Double pendulum The rigid double pendulum impacting a rough hori- zontal surface is shown in Fig. 3. For rods 1 and 2, the following data are given: the masses m1 = m2 = 3 kg, the lengths L1 = L2 = 2 m, the mass moments of inertia with respect to the axis Oz are IC1 = IC2 = 1 kg m2, the inclination angles with respect to the axis Oy are \u03b81 = 20\u25e6 and \u03b81 = 30\u25e6, and the magnitudes of the angular speeds are \u03c91 = 1 and \u03c92 = 2 rad/s. Figure 7 illustrates the energy variation T as function of the coefficient e, for different values of the coefficient s and constant coefficient of kinetic friction, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000756_demped.2007.4393067-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000756_demped.2007.4393067-Figure4-1.png",
+        "caption": "Fig. 4. Schematic diagram summarizing the main steps ofthe digital signal In Fig. 7 the measured curve for an outer race fault of bearprocessing according to the Welch-method ing 1 (motor side) is displayed. The second curve (green line),",
+        "texts": [
+            " This aspect leads to the conclusion that the intensity of the The calculation of the frequency response is carried out by defect, periodically generated at flu, will be modulated by the applying the Welch-method, which is known from the field of mechanical rotation frequency f,,. As modulated signals pro- communications engineering. The main steps for calculating duce spectral components at the sum- and difference- the frequency response of a system G(jc) according to the frequencies, sidebands around flu will appear at Welch-method by using the input data u(k) and the output data fSBRF(v, )= (v - (11) y(k), k =1,2,...,M are briefly sketched in the diagram whichS,(IR ) S,TF is depicted in Fig. 4. More details concerning the digital signalfSB 'r V ) fF = 23 ..Iv=123. 12SRF ( V ) + =1,2,3, (12) processing can be found in the literature [18], [19]. As the The upper index SB stands for sideband. Index r denotes the measurement is carried out in a closed loop control system, right sideband frequency and t the left sideband frequency. two measurement signals are required in order to get reliable The width of these sidebands is results [18]. Therefore, the signal processing algorithms must AfSB = f (13) be applied twice"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000373_50006-8-Figure6.11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000373_50006-8-Figure6.11-1.png",
+        "caption": "Fig. 6.11. Material and Fabrication Savings with Powder Metallurgy I~'",
+        "texts": [
+            " Compared to conventional forging, these processes offer the following advantages: reduced flow stresses, enhanced workability, greater microstructural control and greater dimensional accuracy. The low strain rates, in combination with fine grain sizes, allows production of near net shapes which reduces material and machining cost. HIP followed by forging can be more economical than conventional cast or wrought products, due to improved material utilization, fewer forging steps and reduced final machining. A comparison between conventional ingot metallurgy and forging with the powder metallurgy approach is shown in Fig. 6.11. The biggest disadvantages are high die costs (e.g., TZM molybdenum) and the requirement for forging in vacuum or in an inert atmosphere. Superalloy powders can also be consolidated by extrusion. Powder, normally containerized, is hot extruded at a reduction ratio in the range of 13-1 to achieve a fully consolidated billet. The extruded billet is then hot worked by conventional press forging, isothermal forging or hot die forging. More recently, as-HIP PM products (no forging), which require only minimal machining, are emerging as an even lower cost route"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000804_robot.2007.364174-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000804_robot.2007.364174-Figure2-1.png",
+        "caption": "Fig. 2 Definition of arm angle \u03c8.",
+        "texts": [
+            " With these coordinate systems, the Denavit-Hartenberg parameters are described as listed in Table 1. Note that the notation of the parameters is not unique because the parameters depend on the definition of the joint coordinate frames. Since the tip pose is uniquely described by six parameters, an additional parameter is required to specify the manipulator\u2019s posture uniquely. This parameter is associated with a self-motion of the manipulator. To describe the self-motion, this paper incorporates the parameter termed arm angle [10]. As shown in Fig. 2, the arm angle \u03c8 is defined as the angle between a reference plane and the arm plane spanned by the shoulder, elbow, and wrist. In this paper, the reference plane is determined in the following way. The S-R-S manipulator can be regarded as a non- redundant manipulator when the joint angle 3 is fixed to zero. In this case, the arm plane is uniquely determined for a specified tip pose because no self-motion is possible for the virtual non-redundant manipulator. Since the arm plane is always deterministic except for the case of shoulder/elbow singularity, it can be used as the reference plane",
+            " (4) Since the self-motion parameterized by the arm angle \u03c8 is a rotation around this fixed axis, the self-motion does not affect the wrist position but does affect the wrist orientation. The orientational difference between the arm plane and the reference one is described [11] by 0R\u03c8 = I3 + sin\u03c8 [ 0usw\u00d7 ] + (1 \u2212 cos\u03c8) [ 0usw\u00d7 ]2 , (5) where I3 \u2208 3\u00d73 is the identity matrix, 0usw \u2208 3 is the unit vector of 0xsw, and [ 0usw\u00d7 ] denotes the skew-symmetric matrix of the vector 0usw. Thus, the wrist orientation is described by 0R4 = 0R\u03c8 0Ro 4, (6) where 0Ro 4 represents the wrist orientation when the arm plane coincides with the reference plane. As Fig. 2 implies, the elbow angle (joint 4) depends only on the wrist position and is not subject to the arm angle. It follows that 3R4 and 3Ro 4 are same. Therefore, (6) can be simplified to 0R3 = 0R\u03c8 0Ro 3. (7) Substituting (7) into (1) and (2), we can get the kinematic relation including the arm angle parameter. 2) Inverse Kinematic Analysis: First, the elbow joint angle is computed. Substituting (1) into (4), we have 0xsw = 0R3 ( 3lse + 3R4 4lew ) . (8) Computing the square sum of the norm of 0xsw, we get \u2225\u22250xsw \u2225\u22252 = d2 se + d2 ew + 2dsedew cos \u03b84"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002977_robot.2010.5509531-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002977_robot.2010.5509531-Figure8-1.png",
+        "caption": "Fig. 8. Swing-leg retraction",
+        "texts": [
+            " A symmetric rimless wheel has the following recurrence formula: K\u2212[i + 1] = \u03b5K\u2212[i] + \u0394E, (17) where \u03b5 = K+/K\u2212 [-] is the energy-loss coefficient and \u0394E [J] is the restored mechanical energy [12]. In rimless wheel, \u03b5 and \u0394E are kept constant and the generated gait thus converges to one-period stable limit cycle. In general limit cycle walkers, however, these are not kept constant and thus the limit cycle analysis becomes complicated. Especially, swing-leg retraction (SLR) strongly changes the value of \u03b5. SLR is a phenomenon that the swing-leg moves backward just prior to impact [4] as shown in Fig. 8, and has a great influence on the gait efficiency and limit cycle stability [5][6][7][8]. Let \u03bd := \u03b8\u0307 \u2212 2 /\u03b8\u0307 \u2212 1 [-] be the ratio of angular velocity just before impact, then the angular velocity vector can be arranged as \u03b8\u0307 \u2212 = [ 1 \u03bd ]T \u03b8\u0307 \u2212 1 . The kinetic energies just before and just after impact are also arranged as K\u2212 = 1 2 [ 1 \u03bd ]T M [ 1 \u03bd ] ( \u03b8\u0307 \u2212 1 )2 , (18) K+ = 1 2 [ 1 \u03bd ]T \u039eTM\u039e [ 1 \u03bd ] ( \u03b8\u0307 \u2212 1 )2 . (19) The energy-loss coefficient is then rewritten as \u03b5 = [ 1 \u03bd ]T \u039eTM\u039e [ 1 \u03bd ] [ 1 \u03bd ]T M [ 1 \u03bd ] = N\u03b5 D\u03b5 , (20) where N\u03b5 = I(I + I\u03bd2 + 2ml2) +4ml2 cos(2\u03b1)(I\u03bd + ml2 cos(2\u03b1)), (21) D\u03b5 = (I + 2ml2)(I + I\u03bd2 + 2ml2)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003982_clen.201200075-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003982_clen.201200075-Figure4-1.png",
+        "caption": "Figure 4. Permeation fluxes of the reverse osmosis membranes.",
+        "texts": [
+            " Considering the pre-treatment and final treatment processes together, the highest total COD removal efficiency with the BW30 membrane was found to be 95.7%, while the same was determined to be 96.3% with the XLE membrane. Permeation fluxes were continuously monitored with the aid of a digital balance connected to a computer. Permeate collection line was discharged after each different experiment. Permeation collection line was filled with the new permeate before the determination of new permeation flux. The permeation flux values obtained are shown in Fig. 4. Permeation fluxes increased linearly with transmembrane pressure. Fluxes in the XLE membrane were found out to be higher than those in the BW30 membrane. While applying ultrafiltration as a pre-treatment process caused the BW30 flux value to decrease slightly, it increased the flux values of the XLE membrane. The highest permeation flux values obtained for BW30 in both cases with and without ultrafiltration as the pre-treatment process were 15.3 and 14.6 L m 2 h 1, while the same were 17.5 and 21"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002258_978-1-4419-1126-1_13-Figure13.17-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002258_978-1-4419-1126-1_13-Figure13.17-1.png",
+        "caption": "Fig. 13.17 (a) Direct ophthalmoscopy with Navarro\u2019s schematic eye [22]. (b) Ophthalmoscopy with Navarro\u2019s schematic eye with a vitrectomy lens [23]. (c) Indirect ophthalmoscopy with Navarro\u2019s schematic eye with a condensing lens [63] (#IEEE 2008), reprinted with permission",
+        "texts": [
+            " Throughout this section, the object\u2019s depth z is measured along the optical axis. We begin by investigating the feasibility of imaging and localizing intraocular devices using existing ophthalmoscopy methods. In a relaxed state, the retina is projected through the eye optics as a virtual image at infinity. An imaging system can capture the parallel beams to create an image of the retina. In direct ophthalmoscopy the rays are brought in focus on the observer\u2019s retina [57]. By manipulating the formulas of [56] the field-of-view for direct ophthalmoscopy is found as 10 \u2218 (Fig. 13.17a). Every object inside the eye creates a virtual image. These images approach infinity rapidly as the object approaches the retina. Figure 13.18 (solid line) displays the distance where the virtual image is formed versus different positions of an intraocular object. In order to capture the virtual images that are created from objects close to the retina, an imaging system with near to infinite working distance is required. Such an imaging system will also have a large depth-of-field, and depth information from focus would be insensitive to object position (Table 13.2). To visualize devices operating in the vitreous humor of phakic (i.e. intact intraocular lens) eyes, only plano-concave lenses (Fig. 13.17b) need to be considered [57]. Vitrectomy lenses cause the virtual images of intraocular objects to form inside the eye, allowing the imaging systems to have a reduced working distance. Based on data given from HUCO Vision SA for the vitrectomy lens S5. 7010 [23], we simulated the effects of a plano-concave vitrectomy lens on Navarro\u2019s eye (Fig. 13.17b). This lens allows for a field-of-view of 40 \u2218 , significantly larger than the one obtainable with the method described in Sect. 4.1.1. As shown in Fig. 13.18 (dashed line), the virtual images are formed inside the eye and span a lesser distance. Thus, contrary to direct observation, imaging with an T a b le 1 3 .2 O p ti ca l p ar am et er s fo r th e sy st em s o f F ig .1 3 .1 7 (# IE E E 2 0 0 8 ), re p ri n te d w it h p er m is si o n S u rf ac e 1 2 3 4 5 a 5 b 5 c 6 b 6 c 7 R a d iu s (m m ) 1 2 ",
+            "0 7 T h ic k n es s (m m ) 1 6 .3 2 4 .0 0 3 .0 5 0 .5 5 1 2 .0 0 2 .0 0 1 1 3 .0 0 1 R ef ra ct io n in d ex 1 .3 3 6 1 .4 2 0 1 .3 3 7 1 .3 7 6 1 .0 0 0 1 .4 2 5 1 .0 0 0 1 .0 0 0 1 .5 2 3 1 .0 0 0 optical microscope (relatively short working distance and depth-of-field) is possible. The working distance of such a system must be at least 20mm. As depth-of-field is proportional toworking distance, there is a fundamental limit to the depth-from-focus resolution achievable with vitrectomy lenses. Indirect ophthalmoscopy (Fig. 13.17c) allows for a wider field of the retina to be observed. A condensing lens is placed in front of the patient\u2019s eye, and catches rays emanating from a large retinal area. These rays are focused after the lens, creating an aerial image of the patient\u2019s retina. Condensing lenses compensate for the refractive effects of the eye, and create focused retinal images. We simulated the effects of a double aspheric condensing lens based on information found in [63]. This lens, when placed 5mm from the pupil, allows imaging of the peripheral retina and offers a field-of-view of 100 \u2218 ",
+            " This method was first introduced by Bergeles et al. [12]. As previously stated, the condensing lens projects the spherical surface of the retina onto a flat aerial image. Moving the sensor with respect to the condensing lens focuses the image at different surfaces inside the eye, which we call isofocus surfaces. The locus of intraocular points that are imaged on a single pixel is called an isopixel curve. Figure 13.21a shows a subset of these surfaces and curves and their fits for the system of Fig. 13.17c. The position of an intraocular point is found as the intersection of its corresponding isopixel curve and isofocus surface. The location of the isofocus surfaces and isopixel curves are dependent on the condensing lens and the individual eye. The optical elements of the human eye can be biometrically measured. For example, specular reflection techniques or interferometric methods can be used to measure the cornea [46], and autokeratometry or ultrasonometry can be used to measure the intraocular lens [37]",
+            "22a\u2013d), and thus, 3D intraocular localization with a wide-angle is unambiguous. As an experimental testbed, we use the model eye [31] from Gwb International, Ltd. This eye is equipped with a plano-convex lens of 36mm focal length that mimics the compound optical system of the human eye. Gwb International, Ltd. disclosed the lens\u2019 parameters so that we can perform our simulations. We also measured the model\u2019s retinal depth and shape. The optical system under examination is composed of this model eye and the condensing lens of Fig. 13.17c, where the refraction index was chosen as 1. 531. The simulated isofocus surfaces and isopixel curves of the composite system, together with their fits, are shown in Fig. 13.21b. Based on these simulations, we parametrize the isofocus surfaces and the isopixel curves (Fig. 13.22e\u2013h). The behavior of the parameters is similar to the one displayed in Fig. 13.22a\u2013d for Navarro\u2019s schematic eye. We assume an invariant conic constant of 1. 05, because the variability of the surfaces can be captured sufficiently by the curvature"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000923_bfb0110378-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000923_bfb0110378-Figure1-1.png",
+        "caption": "Fig. 1. torsion of a flexible beam",
+        "texts": [],
+        "surrounding_texts": [
+            "02tO(x,t) = O~O(x,t), x e [0,1] ox0(o, t) = u(t) OxO(1, t) = 0t20(1, t), where O(x,t) is the torsion of the beam and u(t) the control input. From d'Alember t ' s formula , 0(x, t) = r + t) + r - t) , we easily deduce 20(t ,x) = y(t + x - 1) - y(t - x + 1) + y(t + x - 1) + y ( t - x + 1) 2u(t) = 9(t + 1) + ~/(t - 1) - ~)(t + 1) + y(t - 1), where we have set y(t) := 0(1, t). This proves the system is &free with 0(1, t) as a \"&flat\" output. See [46,15,18] for details and an application to motion planning. Many examples of delay systems derived from the 1D-wave equation can be treated via such techniques (see [8] for tank filled with liquid, [14] for the telegraph equation and [57] for two physical examples with delay depending on control)."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003412_978-3-642-28359-8_10-Figure10.8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003412_978-3-642-28359-8_10-Figure10.8-1.png",
+        "caption": "Fig. 10.8 Schematic of a laser deposited track to define \u201ctransverse traverse index (i = x/W )",
+        "texts": [
+            " The effects of these three parameters can actually be accounted with two parameters viz. laser energy per unit transverse length (El = PL/Vs) and powder fed per unit transverse length (m p/ l = m p/Vs). The distance between two adjacent clad tracks, along with above two parameters, can be used in cross thin walled fabrication strategy to generate materials with different porosities. The ratio of the distance between two adjacent tracks (x) and width of each track (W) known as \u201ctransverse traverse index\u201d (i) and is schematically shown in Fig. 10.8. The effect of processing parameters on porosity in laser rapid manufactured structures of Inconel-625 was evaluated by experiments using Box-Behnken array of surface response methodology [31]. The range of processing parameters was: El = 150\u2013300 kJ/m; m p/ l = 16.67\u201336.67 g/m and i = 0.7\u20131.3. The results of the above experiments, with corresponding processing parameters, were subjected to the Analysis of Variance (ANOVA). The calculated standard F-test parameters of ANOVA for m p/ l , i and El were 155"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-155-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-155-1.png",
+        "caption": "Figure 3-155: Leaching.",
+        "texts": [],
+        "surrounding_texts": [
+            "\u2022 Replace any TFE V-rings, wedge rings, or taper rings with elastomer O-ring secondary seals. Elastomer O-ring secondary seals are less susceptible to fretting corrosion because they are better able to absorb minor internal axial shaft movement. \u2022 Replace a pusher-type seal with a nonpusher-type seal such as a metal or elastomer bellows seal having static secondary seals. 3. Chemical Attack on O-rings Failure mode. Swollen O-rings or O-rings that have taken a permanent set preventing axial movement of the sliding seal face are likely to have been damaged by chemical attack. Chemical attack can leave O-rings hardened, bubbled, broken, or blistered on the surface, or with the appearance of having been eaten away. See Figure 3-154. Causes. The most likely causes of chemical attack are incorrect material selection or the loss or contamination of the seal buffer fluid. Action. \u2022 Make a complete chemical analysis of the product and reevaluate the O-ring material selection. Frequently, the presence of trace elements, originally overlooked when specifying the seals, will be at fault. If a suitable material cannot be found, create an artificial environment by flushing the seal from an external source, or use a seal design that incorporates a static secondary seal component. Machinery Component Failure Analysis 229"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002894_kem.450.461-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002894_kem.450.461-Figure1-1.png",
+        "caption": "Fig. 1 Simulation model and raster scanning pattern",
+        "texts": [
+            " A new layer is then activated (via element birth) and the necessary boundary conditions and inputs applied. The elements before they are activated are visually present, however, they do not add to the overall stiffness matrix. The entire process is again repeated until the last layer is complete. A sequential coupled thermo-structural field problem using the transient temperature loads of the thermal analysis as input for the structural model was used for the simulation. To prevent rigid body movement, a zero displacement boundary condition is applied to one side of the platform. Fig. 1 shows the finite element model and scanning pattern. The model consists of two 30m layers of TiAl6V4 lased on a hot-rolled steel platform. The material properties for TiAl6V4 and hot-rolled mild steel (AISI 1015) for the density, thermal conductivity, enthalpy and latent heat of fusion were taken from Mills [7] and Touloukian [8]. The thermal conductivity of the powder which is a field property was taken as 0.25W/mK for temperatures below the melting point [9]. Non-field properties which are dependent on the bulk density of a composition are computed based on measurements",
+            " Typical process parameters for building TiAl6V4 parts were used for the experiment \u2013 the laser power and speed were set to 195W and 1200mm/s respectively with no scan track overlaps [11]. The topography experiment involved capturing the surface state of the platform area before and after the laser melting using an Olympus Confocal Laser Scanning microscope [12]. The platform surface was divided into cells of 480\u00b5m \u00d7 640\u00b5m. A series of images taken at intervals were then knitted to form a 4800\u00b5m \u00d7 5120\u00b5m surface topography plot. Vertical displacements for designated paths A-A and B-B (See Fig. 1) were obtained as the difference between the primary height profiles before and after laser application. The entire process was repeated for three samples, from which the average values were obtained. Fig. 2 shows the surface topographic plots for the experiment. Due to visually evident splattering of molten material in the scanned zone, the displacement changes are measured for the regions outside the scanned zone. Fig. 2 Surface topographic pots before and after laser scanning Fig. 3 compares the surface profile heights from experiments and the simulation model for the paths designated A-A (X-direction) and B-B (Y-direction)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003968_cjme.2012.05.968-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003968_cjme.2012.05.968-Figure6-1.png",
+        "caption": "Fig. 6. Mixed model in ADAMS",
+        "texts": [
+            " In this paper, we build the outline curve model according to the second method. A typical model is shown in Fig. 5. Establish the outline curve model of each part involved in contact according to the method mentioned above, insert them into the solid model respectively and adjust the location of them to coincide with the outline of the solid models. Then, determine a contacting plane for each pair of the contacting parts and move the curve models onto the corresponding planes. So far, a completed mixed model is established. An example of the mixed model is shown in Fig. 6. Fig. 7 shows the schematic diagram of runaway escapement, consisting of two stages of gear and escapement. The transmission of motion or power among parts is achieved by the contact constraints. There are four contact constraints in the whole system. Apply a constant torque (T0.1 N \u2022 mm) on gear A, choose GSTIFF as the integrator of dynamics model, set the calculation error of 0.000 1 and the minimum time step of 0.000 1 s, solve the solid model and the mixed model in the duration time of 1.4 s respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001925_j.jelechem.2009.10.022-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001925_j.jelechem.2009.10.022-Figure1-1.png",
+        "caption": "Fig. 1. Schematic illustration of the SWNT-sheet modified working electrode.",
+        "texts": [
+            ", Tokyo, Japan) were used for the morphological observations. The SWNT-sheet modified electrode was prepared as followings: A glassy carbon electrode (3 mm diameter) was polished with chamois leather containing Al2O3 slurry and then ultrasonically cleaned in milli-Q water and wiped with filter paper. The SWNTsheet was then directly covered onto the dry surface of the glassy carbon electrode, and immobilized with plastic tape which was opened in the center of electrode surface around 3 mm in diameter as shown in Fig. 1. The SWNT cast electrode was also prepared as a working electrode. Briefly, a glassy carbon electrode (GCE) was polished with chamois leather containing Al2O3 slurry and then ultrasonically cleaned in milli-Q water and wiped with filter paper. About 5 lL of aqueous suspension containing 2 wt.% of the dispersed SWNTs and 1.0 sodium dodecyl sulfate was dropped on the GCE surface and then dried at room temperature to evaporate the aqueous solvent. The SWNT suspension was repeatedly dropped on GCE surface in order to increase the contents of SWNTs on the electrode"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001274_20080706-5-kr-1001.02709-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001274_20080706-5-kr-1001.02709-Figure1-1.png",
+        "caption": "Fig. 1. Variable definition for kinematic, 2D UAV, bird\u2019seye view with z-axis pointing outwards from the paper.",
+        "texts": [
+            " Nearest neighbor rules, Lyapunov theory, graph techniques and non-smooth control results are used to study how multiple agents eventually move in the same direction despite the absence of centralized coordination and despite the fact that each agent\u2019s set of nearest neighbors changes with time as the system evolves (Tanner et al. [2003a,b]). We consider the design of mesh-stable controllers for vehicle formations. For the sake of example, we will consider a simple model for small fixed-wing UAVs operating in a 2-D plane. A related, mesh-stable approach for helicopters can be found in (Pant et al. [2001]). A design for fixed wing aircraft can be found in (Spry and Hedrick [2004]). In our development, we have used a kinematic model to represent the aircraft. The reference frames are shown in Fig. 1.    x\u0307 = u1 cos\u03c8 + Vwx y\u0307 = u1 sin\u03c8 + Vwy \u03c8\u0307 = u2 (1) Here, Vwx and Vwy represent the velocity due to wind in both x and y directions. The directions x and y are in a frame that is fixed with respect to the ground. The control variables are u1 and u2, the airspeed and turn rate. In the case of model aircraft it is common to keep the airspeed u1 within a fairly small range and use the turn rate u2 as the primary control variable. We are restricting ourselves to aircraft moving in the horizontal plane for now"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure4.28-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure4.28-1.png",
+        "caption": "Figure 4.28 Angle of departure example",
+        "texts": [
+            " The other roots along the real axis loci correspond to that same value of loop gain. Rule 5 When two adjacent poles lie on the real axis, there will be a breakaway point on the locus between these two poles. These breakaway points are defined: Rule 6 The angle of departure of a locus from a complex pole can be determined from the following equation, where is the angle of departure: = 180\u2212\u2211 p +\u2211 z where \u2211 p = \u2211 (pole angles to the pole), and \u2211 z =\u2211 (zero angles to the pole). 146 Introduction to Laplace Transforms Figure 4.28 illustrates this rule for a system with two complex conjugate poles, a real pole and a real zero. The above six rules of root locus construction make it easy to generate the root loci for any linear system and to visualize how the closed loop roots move as the loop gain is varied from zero to infinity. Before applying these rules to a specific control system example we need to have a feel for how root locus and the \u2018s\u2019 plane relate to the fundamentals of control systems analysis already covered"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001141_j.jmatprotec.2008.02.071-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001141_j.jmatprotec.2008.02.071-Figure12-1.png",
+        "caption": "Fig. 12 \u2013 Pressure angles at different points.",
+        "texts": [
+            " As vectors TB and T\u2032 B are perpendicular to \u2212\u2192 BO and \u2212\u2192 BG, respectively, the relief angle B is obtained as B = 2 \u2212 B (53) 4.2.3. Analysis of the clearance angle E In Fig. 9, TE is the tangential vector at point E while T\u2032 E is the vector perpendicular to the position vector rE. Applying the theory of involutometry, the clearance angle E is obtained as 854 j o u r n a l o f m a t e r i a l s p r o c e s s i n g t e c h n o l o g y 2 0 9 ( 2 0 0 9 ) 847\u2013855 and the pitch circles of the generated gears, the arc length of NJ shown in Fig. 12 is equal to the corresponding distance changed according to the shifted amount. follows: E = 2 + \u2212 E (54) where = inv B \u2212 inv E (55) 4.2.4. Effect of shifted distance on the cutting angles By substituting the values from Table 1, the various cutting angles are obtained and shown in Fig. 10 where the outside diameters are changed according to the shifted distance. If the outside diameters are kept constant, the cutting angles are as shown in Fig. 11. kept constant. 4.3. Width of top land of spur-typed cutter The width of the top land of the spur-typed cutter can be obtained by solving for the coordinates of points A and B shown in Fig. 12. The coordinates of point A were solved in the previous section. If the polar angle of point B is solved, the x-component of point B is Bx = rB cos and the y-component is By = rB sin . From the geometric relationship shown in Fig. 12, can be expressed as = \u2220KOE + \u2220EOB (56) where \u2220EOB = inv B \u2212 inv E (57) Due to pure rolling between the pitch line of the rack cutter j o u r n a l o f m a t e r i a l s p r o c e s s i n g t e c s \u0131 w t F t t d 5 A t c t a c r onstant outside diameter, and variable outside diameter ccording to the shifted amount, respectively. hown in Fig. 13. The angle \u0131=\u2220NOJ is therefore = r1 (58) here r1 is the radius of the pitch circle of the generated spuryped cutter. Therefore, angle is obtained as = inv B \u2212 inv J + \u0131 (59) ig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure4-1.png",
+        "caption": "Fig. 4 Generating process using tw erating circular-arc rack: double-con",
+        "texts": [
+            " 132, MARCH 2010 om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201 can be solved by Eqs. 11 and 12 . Likewise, the work gear surface equations can be solved based on system equations Eqs. 6 \u2013 12 . The generating cutting process of the longitudinal cycloidal helical gear is commonly explained by an imaginary virtual generating circular-arc rack formed by the locus of the cutter blades. In this process, all inside cutter blades mounted on the head cutter are used for a double-convex gear see Fig. 4 , and all outside cutter blades are used for a double-concave gear see Fig. 5 . 5 Simulation of the Condition of Meshing and Contact The TCA technique has also been applied to simulate the condition of meshing, whose main goal is to determine the shift of the bearing contact and the transmission errors caused by the misalignment of the gear drive. The first step in such application is to represent the equations for two mating tooth surfaces in the same coordinate system, whose coordinates and unit normal should be the same at the point of contact"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000647_tmag.2007.893803-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000647_tmag.2007.893803-Figure2-1.png",
+        "caption": "Fig. 2. Simplified analytical model with only PM field (A) when stator and rotor shaft are air and (B) considering the stator and rotor shaft core.",
+        "texts": [
+            " The analytical results are validated by comparison with finite element analyses and experimental results. Fig. 1 shows the 2-D model of a typical 4-pole, 3-phase PM synchronous motor equipped with a parallel magnetized PM rotor and a slotted stator core. The main design specifications of PM synchronous motor are given in Table I. In order to establish analytical solutions for the magnetic field distribution, this paper assumes that the relative permeability of the stator core and rotor shaft is infinite and the current is distributed in an infinitesimal thin sheet at . Fig. 2 shows the simplified analytical model with only PM field for deriving transfer relations at the magnet surface [(d) and (e) in Fig. 2(A)]. Since there is no free current in the PM region [4] (1) Digital Object Identifier 10.1109/TMAG.2007.893803 The magnetic vector potential A is defined as . By the geometry of the cylindrical structure, the magnetic vector potential has only which is independent z. In the polar coordinate system, the magnetization M is given by (2) In nonconducting regions, the magnetic vector potential is assumed to have Coulomb gauge dependence and satisfies the Poisson equation (3) (4) 0018-9464/$25.00 \u00a9 2007 IEEE where denotes the free space permeability, is the spatial wave number of the th harmonic. Here, is the complex Fourier coefficient of th order parallel magnetization component (5) By , flux density of normal and tangential component could be expressed as (6) Using general solution of (4), the transfer relation between the vector potential and the flux density which relate the field evaluated at the magnet region\u2014(d), (e) identified in the model of Fig. 2 are (7) where the geometric parameters are (8) C. Magnetic Field Solutions by PM Regions Fig. 2(B) shows the simplified analytical model for predicting generalized magnetic vector potential due to PM. Since there is no source term in the airgap region, the transfer relations can be easily expressed as follows [elimination only the source term of (7)]: (9) The boundary conditions used in analytical prediction of the magnetic vector potential due to the PMs are as follows: and Following an analysis similar to (9), the transfer relations at an imaginary boundary in airgap regions can be expressed as (10) inary boundary "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003086_pime_conf_1966_181_311_02-Figure8.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003086_pime_conf_1966_181_311_02-Figure8.1-1.png",
+        "caption": "Fig. 8.1. Rigid cylinder and plane",
+        "texts": [],
+        "surrounding_texts": [
+            "The equation to be solved is the well-known Reynolds equation, Proc Instn Mech Engrs 196647 The particular form of the Reynolds equation presented above is based upon the assumption that the lubricant is incompressible and isoviscous. The Martin theory The solution for a rigid cylinder of infinite extent has been presented by Martin (I). In this instance side leakage can be neglected, and the equation reduces to (A3g) = 1 2 7 2 4 ~ ah . . (8.2) This equation can be integrated to give h-h, dx (8.3) If a rigid circular cylinder near a plane is considered (see Fig. S.l), then R n h = --(l+ncos 8) . . (8.4) where n = -s/ l+s and s = R/h,. When this film thickness expression is introduced into equation (8.3) the Sommerfeld variable can be employed and an integration performed to yield the pressure distribution in terms of 0: p = 0 when 8 = Or (inlet) p = - aP = 0 when 0 = O2 (outlet) ax Since the effect of the selection of 8, is negligible so long as 8, is not too small, it is frequently taken as -90\". This point has been fully considered by Dowson and Whitaker (4- Vo1181 Pt 3 0 at WEST VIRGINA UNIV on June 5, 2016pcp.sagepub.comDownloaded from SIDE-LEAKAGE FACTORS FOR A RIGID CYLINDER LUBRICATED BY AN ISOVISCOUS FLUID 167 Normal and tangential surface force components can be defined as follows : Px, = 0 02 PXz = -I, pR sin 0 d8 F1 = [ T , = ~ R cos 9 dO+ T , = ~ fR cos 8 d0 l2 F, = -[r,=,,RcosOdO- r ry=,,fRcos8d0 where f = h,/h. It can further be shown that . (8.6) ! F P X Z 1 - 2 7(U,-Uz)A P X Z Fz = --+7(U,-U2)A 2 Solutions have been presented for these force components for a wide range of Rih, in reference (3). For R/ho greater than lo4 a good approximation to the force components per unit axial length is given by i Px, = 0 1 (8.7) I Short-cylinder theory Following Dubois and Ocvirk (4) it is assumed that the first term in equation (8.1) can be neglected compared with the second term if the axial length of the cylinder is very small compared with the radius. It will be seen later that the cylinder has, in fact, to be very short indeed for this condition to apply. The Reynolds equation reduces to (h3:) = 12724- ah ax . . (8.8) Since h is a function of x only, this equation can be integrated twice to yield the pressure distribution. With the boundary conditions that p = 0 when z = &l, the solution becomes 6724 ah p = - - ( 2 2 - 1 2 ) . . (8.9) h3 ax Difficulties are encountered if a circular profile is considered, since ahlax and hence p achieve an infinite value when 0 = -90\". However, if a parabolic profile given by X 2 2R h = ho+- . . . (8.10) is considered, the difficulty is removed. Proc Instn Mech Engrs 1966-67 12 It can be seen from equation (8.9) that the pressure returns to the ambient value along the line of closest approach, and if the sub-ambient pressures predicted for the divergent clearance space are neglected a normal force component can be defined as P, = j:lj:apdxdz . . (8.11) The solution is 13 p y = 2 v ( U i + U 2 ) ~ . . (8.12) The comparable solution for a width 21 of an infinitely long cylinder, as given by the Martin theory, is P, = 2*45q(U1+ Uz)2Z - . (8.13) (a An important difference between the two solutions is that the 'short-cylinder' load capacity is independent of R. It will be noted that the two solutions predict the same load capacity when I . . . -=JE0 R R (8.14) Full solution The Martin and short-cylinder solutions represent extremes which are clearly very different and which are likely to lead to erroneous predictions in the important intermediate range of cylinder sizes. A 'full' solution of equation (8.1) has been obtained numerically. For this analysis it is convenient to write equation (8.1) in dimensionless form, as follows : 2 ax (H3$)+(:)z$(H3g) =g (8.15) A where x = , + 1 h H=h, Pho2 p = - 127uR It is well known from the two-dimensional (Martin) case that the pressure distribution exhibits a very localized pressure field giving high values of apiax. Such a condition is not favourable to the relaxation process. In order to produce a much gentler curve, a parameter 4 used by Vogelpohl (5) is introduced which has a value PH312; P being small when His large, and vice versa. This substitution also has the advantage of eliminating all terms containing derivatives of products, or products of derivatives, of H and P or H and 4. The equation to be solved is now . . . (8.16) Yo1181 Pr 3 0 at WEST VIRGINA UNIV on June 5, 2016pcp.sagepub.comDownloaded from 168 D. DOWSON AND T. L. WHOMES In order to solve equation (8.16) by relaxation methods it is necessary to introduce finite-difference representation. A mesh is set up having equal increments in the axial ( z ) direction and variable steps in the direction of motion (x) as shown in Fig. 8.2. The variable steps are necessary to establish adequate accuracy of representation of the rapidly varying functions in the x-direction. If, at three points x3, xo, xl, a function of x, say g, is represented by a quadratic, g = Ax\"+Bx+C and the intervals between x3, xo and xo, x1 are e and f respectively, the constants A, By C take the following form when the origin is located at x3 : B = - e\"1 - g 3 f P + f ) +go(e+f)\" ef @+f) c = g3 The derivatives in equation (8.16) can now be expressed in finite-difference form in terms of 4 and H a t surrounding stations as follows : i3H e2Hl -f2H3 + (f\" - e2)Ho _- ax - ef (e+f I a ax = {e2Hl[H11'2(e2+2ef) +Ho1/2(f2 -2) + H31'2f2] + Ho[- H11'2e2(e+f)2+ Holi2(f2-e2) +f2H3[H11i2e2+ Ho1'2(e2-f2) - H3\"2f2(e +f)\"I + H31'2(2ef+f2)l)/e2f2(e+f)2 1cI - _ - p 2 say The latter finite difference representation is formed by fitting a parabola to the values of H1'\"(aH/aX) at the Proc Instn Mech Engrs 1966-67 b2{e2Hl - f 2 H 3 + ( f 2 -e2)Ho} Ho3I2p and G = (A) The following boundary conditions are adopted : At inlet p = 0 when x = x1 At the edges p = 0 when z = f l aP aP ax az At outlet p = - = - = 0 The latter condition represents the usual cavitation boundary condition and it has been considered in detail by Jakobsson and Floberg (6). The normal and tangential force components can be evaluated by integration of the solutions for pressure distribution. Thus P, = 2 / ' r r n p d x d z . . (8.18) 0 x1 where xm is the value of x at the cavitation boundary. Similarly, Px = 2 / ' r p t a n B d x d z . (8.19) 0 x1 and F = 2 l/Iy T dx dz+2 1' [ ZT dx dz (8.20) 0 Xrn where T is the viscous shear stress on the surface of the solid and f (= hm/h) represents the proportion of the cylinder width occupied by lubricant in the cavitated zone. COMPUTATION A solution to equations (8.16)-(8.20) was obtained by means of a computer programme, written in Algol 60 and run on an English Electric K.D.F.9 high-speed digital computer. The position of the mesh points was put into the computer as data, such that a fine mesh was obtained where the pressure gradient was expected to be high and the cavitation boundary was expected to occur. The mesh lines in the x-direction were numbered 0 to a from the inlet to the outlet edge, the relaxation being performed for the lines 1 to r where the line r was beyond the expected position of the cavitation boundary. Vol181 Pt 3 0 at WEST VIRGINA UNIV on June 5, 2016pcp.sagepub.comDownloaded from SIDE-LEAKAGE FACTORS FOR A RIGID CYLINDER LUBRICATED BY AN ISOVISCOUS FLUID 169 Since the cylinder was symmetrical about the central plane (z = 0), only half the cylinder was considered, the lines in the axial direction being numbered 0 to q from the edge to the centre-line. The total number of points to be relaxed was, therefore, r x q which was called N. Having calculated the film thickness H on each line in the X-direction, the coefficients in equation (8.17) may be evaluated, and the equation is written for each point in the order (1, l), (2, l), etc., to (r, l), (1,2), (2,2) to (r, 2), etc., to (r, 4) . These equations may be written A+ = B where A is an N x N matrix and 4 and B are N-dimensional vectors. On inspection of A, it was found that nonzero elements occur only on five diagonals. Therefore A may be written as a 5 x N matrix to save computer space. The transformation A = L+D+U was used where L, D, and U are lower diagonal, leading diagonal and upper diagonal matrices respectively. The iteration equation is now [see (7) where K is the iterant. A computing procedure was written to perform this relaxation. Initially 4 was set to any value 2, usually zero. Then, using n as a counter, for each value of n, i.e. each mesh point, 4[n] was set equal to B[n] and from this was subtracted the values of + at the surrounding points multiplied by the corresponding elements from A . This result was then divided by the element of A on the leading diagonal to give the new value of q5[n]. This was performed for n from 1 to N so that 4 was now a new vector. The values of q5 were compared with those before the iteration, and if each value had changed by less than a preset tolerance, the relaxation was deemed to have converged. Otherwise the procedure was repeated, In order to speed convergence an over-relaxation factor was used. Since the relaxation continued beyond the cavitation boundary, the solution would exhibit negative values of q5 4\u2018\u201d\u2019\u201d = D-l[B-(L+ u)+(k\u2019] Prac Instn Mech Engrs 1966-67 in this region, and the boundary condition would not be satisfied as shown in Fig. 8 . 3 ~ . Therefore, a mechanism was included in the programme to find this boundary automatically. After each iteration each negative value of + was set to zero. This had the effect of raising the level of C) in this region, so that in the next iteration the pressures at the points surrounding the one being calculated increased its value slightly until finally no negative pressures were obtained and the boundary condition was satisfied, as shown in Fig. 8.3. It is appreciated that the Reynolds equation gives a solution which would follow the dotted line in Fig. 8.3b if permitted to do so. However, so long as the distance from the cavitation line to the last mesh line is not too great and the mesh is fine in this region, the effect of holding pressures beyond the cavitation line to zero will be small. Two procedures were written to integrate a function over a set of evenly and unevenly spaced points. These were used to calculate the force components from the pressure distribution."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000837_6.2008-7192-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000837_6.2008-7192-Figure7-1.png",
+        "caption": "Figure 7. AEI sensor suite",
+        "texts": [
+            " The control surface rate is the same for all of the surfaces, so control weighting are experimentally determined that cause all control rates to be approximately equal 10-4 10-2 100 102 -100 -50 0 50 100 Frequency, rad/sec M ag ni tu de , d B 10-4 10-2 100 102 -200 -150 -100 -50 0 Frequency, rad/sec P ha se , d eg American Institute of Aeronautics and Astronautics 092407 6 and on the rate limit for the large gust. This uses the maximum control power for the gust maneuver and delivers the best performance, assuming that the control robustness criterion has been met. The SensorCraft wing was instrumented with a variety of sensors as shown if Fig. 7 to allow identification of rigid body and flexible modes. The rotation at the bearing that supported the wing was measured to identify pitch attitude. String pots were used to measure height in the tunnel. Pitch rate sensors were located on the body of the aircraft to get a basis for rigid body pitch rate and an additional pitch rate sensor was located on the wing spar to get a measure of pitch rate contributions from the structural modes. Nz sensors were similarly located on both the body During the altitude hold maneuver in a gust the aircraft must pitch down quickly to maintain altitude which drives a high speed response from the pitch attitude control"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000521_s11012-006-9049-z-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000521_s11012-006-9049-z-Figure3-1.png",
+        "caption": "Fig. 3 Schematic diagram of preliminary apparatus to illustrate how the gyroscope may be tilted upwards to gain potential energy simply by applying a small horizontal force along the axis, in the direction of normal precession due to gravity",
+        "texts": [
+            " The load-cell and amplifier were commercially manufactured. Linearity is better than 1% up to 20 kg load. Its response is very insensitive to lateral bending moments. The vertical pointer serves to show that the load-cell remains horizontal while the precessing gyroscope is being lifted. The experimental technique involves displacing the axle pivot from the central position, as proposed by Laithwaite. Preliminary tests were done to establish the general technique, first of all with a minimum of human intervention. As shown in Fig. 3, the spinning wheel axle was attached through the universal-joint and load-cell to a horizontal arm which was capable of rotating horizontally about a vertical axis OZ. The normal precession induced by gravity was satisfied by rotating this horizontal arm slowly around OZ so as to keep the wheel axle horizontal and perpendicular to the arm. Subsequently, by applying a small pulling force to this arm along the wheel axle in the direction of motion so as to increase the precession velocity, the wheel was caused to tilt upwards"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003276_09596518jsce1018-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003276_09596518jsce1018-Figure3-1.png",
+        "caption": "Fig. 3 Experimental device",
+        "texts": [
+            " That means, if the current in the coil is 0, the control output voltage that is computed by the GSMC algorithm will be infinite. This will cause the instability of the system. Usually in the initial state of the control process, the coil current is 0. In this paper, to avoid the instability problem caused by the 0 current, we choose to output a proper voltage V0 in the coil for a moment before the control process so that when the control algorithm is employed, the coil current is not 0. In this study, V0 5 6(V). The whole experimental device is shown in Fig. 3. In the magnetic force model equation (3), the parameters k and A need to be identified. In order to identify these parameters the maglev system is stabilized by a proportional\u2013integral\u2013derivative (PID) controller. Then the ball is set to be levitated in different positions and the coil current will be collected. Table 2 shows the PID controlled ball in different positions with different coil currents From Table 2, the data can be fitted as a first- power equation i~{0:1048xz0:6043 \u00f051\u00de Since the magnetic force is equal to the gravity of the steel ball when the ball is levitated by the PID controller, it becomes k i H{x 2 ~mg \u00f052\u00de The parameters of k and H can be calculated by equations (51) and (52) knominal~22:3361 \u00f0Nmm2 A2 H 5 5"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002031_s10008-010-1147-0-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002031_s10008-010-1147-0-Figure7-1.png",
+        "caption": "Fig 7 A scheme for charge transport in HRP/CdSe/MS film on OTE",
+        "texts": [
+            " CdSe h\u00fe \u00fe e \u00f0 \u00de \u00f04\u00de Concerning the higher sensitive response to H2O2, therefore, we proposed that the CdSe/MS electron\u2013hole pairs generated under UV light can facilitate the reduction processes in the peroxidase\u2013electrode system. In the presence of electron acceptors (H2O2) in solution, it could deplete the photogenerated electrons locating on the electrode and enhance the electron\u2013hole separation of the charge transfer complex to some extent, leading to the enhancement of the steady-state photocurrent [29]. Figure 7 shows the possible charge transport process in HRP/CdSe/MS film on optically transparent electrode (OTE). When CdSe/MS composite film was used as a matrix of the peroxidase\u2013electrode system, the electrons photogenerated from CdSe/MS composite and the electrons provided by electrode could occur simultaneously to catalyze the reduction of H2O2. The holes photogenerated by CdSe are trapped at the interface between CdSe and pore walls of MS film. The result demonstrates that the HRP/CdSe/MS composite film may further be applied as an HRP biosensor for H2O2 monitoring under UV illumination"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001933_s11071-010-9720-8-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001933_s11071-010-9720-8-Figure3-1.png",
+        "caption": "Fig. 3 Model of force diagram for pinion and gear",
+        "texts": [
+            " The oilfilm supporting force is dependent on the integrated action of hydrodynamic pressure and hydrostatic pressure of HSFD. Figure 2 represents the cross section of HSFD. The structure of this kind bearing should be popularized to consist of 2N (N = 2,3,4, . . .) hydrostatic chambers and 2N hydrodynamic regions. In this study, oil pressure distribution model in the HSFD is proposed to integrate the pressure distribution of dynamic pressure region and static pressure region as described in Sect. 2.1. Figure 3 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; 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; K is the stiffness coefficient of the gear mesh, C is the damping coefficient of the gear mesh, e is the static transmission error and varies as a function of time. 2.1 The instant oil-film supporting force for HSFD To analyze the pressure distribution, the Reynolds equation for constant lubricant properties and noncompressibility should be assumed, then the Reynolds equation is as follows [17] 1 R2 \u2202 \u2202\u03b8 ( h3 \u2202p \u2202\u03b8 ) + 1 R2 \u2202 \u2202z ( h3 \u2202p \u2202z ) = \u221212\u03bc\u03a9 \u2202h \u2202\u03b8 + 12\u03bc \u2202h \u2202t "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000391_9780470061565.hbb065-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000391_9780470061565.hbb065-Figure6-1.png",
+        "caption": "Figure 6. Hybrid biosensor principle: a combination of two sensing principles of thermal and electrochemical measurement.",
+        "texts": [
+            " The feasibility of the system for simultaneous multianalyte determination has been exemplified by glucose and urea or penicillin and urea measurement with a linear range of up to 20 mM for urea, 8 mM for glucose, and 40 mM for penicillin.24 Determinations of three analytes (glucose, urea, and penicillin) and four different analytes (glucose, lactate, urea, and penicillin) in sample mixtures were also demonstrated using similar thermal biosensor arrays.32,33 Another field under current investigation is hybrid sensors, which combine two different measurement technologies into a hybrid (Figure 6). Conventional biosensors are usually classified into categories, such as electrochemical, optical, and thermal sensors according to the detection principle. Each type of sensor has its own merits and drawbacks. Creating hybrid biosensors by combining different detection principles could possibly retain the original advantages and avoid the disadvantages of the respective type of sensor. Furthermore, hybridization of biosensors could also create unique properties. One example of a hybrid biosensor was fabricated and demonstrated by Xie and coworkers (Figure 7)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001033_j.jmatprotec.2007.11.070-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001033_j.jmatprotec.2007.11.070-Figure2-1.png",
+        "caption": "Fig. 2 \u2013 Tool for ploughing-extrusion.",
+        "texts": [
+            " The workpiece is fixed on the working platorm, the tool ploughs the workpiece along the direction of -axis, leaving a finned micro-groove behind. A group of fins re created in one plough stroke. Then the workpiece feeds long the direction of X-axis, and the tool begins to plough ext group of fins. Meanwhile the tool extrudes previously ormed fins in order that the fins can grow higher. To avoid ulging, one end of the strip is fixed, while the other end can xtend and shrink freely along the axis X. The structure of the tool is shown in Fig. 2. The high speed teel whose dimension is 2 mm \u00d7 10 mm \u00d7 200 mm is grinded o the ploughing edge, the major extrusion face Ar, the auxil- iary extrusion face A\u2032 r, the major forming face A , the auxiliary forming face A\u02c7 and the tool flank A\u02db. The front end is the ploughing edge which ploughs the metal and makes the metal flow along the major extrusion face and the minor one. The major extrusion face A generally extrudes and ploughs the split metal and makes metal flows along the major forming face A , and then the metal will form the grooves under the extruding of major forming face A "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001303_j.electacta.2009.07.069-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001303_j.electacta.2009.07.069-Figure3-1.png",
+        "caption": "Fig. 3. RDE voltammograms recorded for the catalytic oxidation of 1 mM GSH at a rotating PB film electrode in pH 1.7 solution. Electrode rotation rates (rpm) are denoted in the figure. Dotted line indicates the RDE voltammogram of a rotating PB film electrode (rpm = 350\u20133650) in pure pH 1.7 electrolyte solution ( = 10 mV s\u22121).",
+        "texts": [
+            " This implies that the heterogeneous oxidation of GSH on PB film electrode is not a fast reaction at high scan rates. Therefore, the rate determining step for GSH oxidation on a PB film electrode is thought to be the chemical reaction 7376 S.M. Senthil Kumar, K. Chandrasekara Pillai / Electrochimica Acta 54 (2009) 7374\u20137381 F n at a v e ( ). b t c f a 3 s ( F l ( i t ( ( s a a u p i 4 a p b c r t 3 o f 10 mV s in pH 1.7 solution (0.02 M HCl + 0.1 M KCl). In all these experiments, freshly prepared PB film electrode was used to record the RDE voltammogram at each rpm. It can be noticed in Fig. 3 that there is a significant increase in oxidation current with the first rpm, clearly indicating that the PB ig. 2. Scan rate dependence of peak parameters for 1 mM GSH catalytic oxidatio ersus 1/2; (C) ECV cat versus log( ) and (D) current function (iCV cat/ 1/2) versus scan rat etween the surface BG species and the organic species, rather than he electrochemical formation of BG from PB under these scan rate onditions. The slow chemical reaction at high scan rates is further seen rom the fact that the i\u2013E trace of the catalytic oxidation process ppears as a peak only up to 100 mV s\u22121, becomes wave-like up to 00 mV s\u22121 and beyond this it shows an exponential increase with can rate (not shown here)",
+            " ECV cat versus log( ) is linear with slope [\u2202ECV cat/\u2202log( )] = 0 mV decade\u22121. This is another characteristic of an ECcat mechnism [29\u201332]. As the shift is not too large, the existence of a artial kinetic (chemical reaction) limitation [29\u201332] in the reaction etween the BG and GSH can be inferred. The scan rate normalized atalytic peak currents, i.e., current function (iCV cat/ 1/2) versus scan ate ( ) exhibits a decreasing trend as shown in Fig. 2D and it is a ypical behavior of ECcat mechanism. .1.2. Rotating-disc electrode investigations Fig. 3 shows a set of RDE voltammograms for the catalysed xidation of 1 mM GSH recorded at a PB film electrode at diferent rotation speeds. These voltammograms were recorded in stationary PB film electrode in pH 1.7 solution: (A) log (iCV cat) versus log( ); (B) iCV cat the potential range 300\u20131000 mV by sweeping the potential at \u22121 S.M. Senthil Kumar, K. Chandrasekara Pillai / Electrochimica Acta 54 (2009) 7374\u20137381 7377 F us \u03c9\u22121 i fi a r o t s i c o e s d o t w c c r t o m m t f l t m h w w a N ig. 4",
+            " The decrease in the magnitude of catalytic current after the maxmum diffusion current was varyingly attributed to a change in the hemical nature or structure of the catalyst during the oxidation f the substrate [34\u201336], an inhibition of the catalytic reaction [37] tc. Alternatively, the adsorption of the substrate on lower oxidised tate of the catalyst for its reaction with the adjacent higher oxiised state as suggested by Lin and Wen [38] for benzyl alcohol xidation on RuO2 electrode. In the present work, what ever may be he origin for the current peak, this is not a reversible process since hen the scan is reversed the forward i\u2013E curve is not retraced, as an be seen from the sample CV for 350 rpm in Fig. 3. Fig. 4(A) shows Levich plots (iRDE cat versus \u03c91/2) for various conentrations of GSH. The limiting current is independent of electrode otation speed right from the initial rpm = 350 for all the concenrations of GSH studied ca. 0.5\u20139 mM, signifying that the catalytic xidation of GSH at the PB film electrode is free from solution ass transfer control. These results reflect kinetic limitations on the ediated GSH oxidation current (iRDE cat ) even at the lower concenrations studied. The Koutecky\u2013Levich plots ((iRDE cat ) \u22121 versus \u03c9\u22121/2) or the data in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003578_j.cad.2011.03.001-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003578_j.cad.2011.03.001-Figure5-1.png",
+        "caption": "Fig. 5. The number and distribution of the tangent points. Each medial point is assigned a label of form pN n , the subscript n indicates the dimension of the programming problem, and the superscript N denotes the number of the tangent points.",
+        "texts": [
+            " (a) The optimal conditions of one-dimensional saddle point pro- gramming are: (i) the number of the characteristic points is n+1 = 2; and (ii) if u is a constant nonzero value, ai = [sin\u03d5i], it means that the two characteristic points are located on the e2. Considering Eq. (21), the components should abide the following inequality according to Eq. (21) sin\u03d51 sin\u03d52 < 0. (22) According to the one-to-one correspondence between the characteristic points and the tangent points and Eq. (22), we know that the two characteristic points or tangent points should lie on the opposite sides of axis e1. In this case, the optimal conditions can be shortened as \u2018\u2018opposite sides criterion\u2019\u2019, as shown in Fig. 5(a). (b) The optimal conditions of two-dimensional saddle point programming are: (i) the number of the characteristic points is n + 1 = 3; and (ii) the eigenvectors ai listed in Eq. (20) are two-dimensional vectors, thus the characteristic points lie in a unit circle centered at the origin and the triangle formed by themmust be an acute triangle to contain the origin. So, the optimal conditions can be shortened as \u2018\u2018acute triangle criterion\u2019\u2019, as shown in Fig. 5(b). In a similar fashion, the components of the eigenvectors should satisfy the following inequality cos \u03d52 \u2212 \u03d51 2 cos \u03d53 \u2212 \u03d51 2 cos \u03d53 \u2212 \u03d52 2 < 0. (23) (c) Degenerate situations: in some special cases, the above optimal conditions of the one- or two-dimensional problem will degenerate. This will lead to the reduction of the number of the tangent points, and the distribution of the tangent points becomes particular. As to the one-dimensional case, the two tangent points will come into coincidence and lie on axis e1. The tangent point corresponds to the end point of MA, as shown in Fig. 5(c). As to the two-dimensional case, two of the associated three tangent points will come into coincidence. The line segment between this point and the third tangent point will correspond to a diameter of the medial axis disk, as shown in Fig. 5(d). (d) Redundant situations: on the contrary to the degenerate situations, if the number of the tangent points between the medial axis disk and the boundary is more than that in the regular cases, this will lead to the redundant case. For the onedimensional case, the number of the tangent points is greater than two, and all the tangent points lie within a semicircle, as shown in Fig. 5(e). Likewise, for the two-dimensional case, the number of tangent points is greater than three, and all the points are not within a semicircle, as shown in Fig. 5(f). According to the saddle point properties and optimal conditions of the MAT, the medial axis points can be classified into two types: (1) Regular points. The conditional saddle points and the de- generate cases of the saddle points, where there are two tangent points between the medial axis disk and the boundary curves. (2) Irregular points. The irregular points include: (i) End point, which corresponds to the degenerate case of the conditional saddle point, where the medial disks have only one tangent point with the boundary"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003443_isp-2012-0075-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003443_isp-2012-0075-Figure3-1.png",
+        "caption": "Fig. 3. Sign convention for error measures (\u03b8e k , De k ).",
+        "texts": [
+            " However, for low rudder rates, there is a possibility that this \u2018switching off\u2019 causes delays in new commands in case ship movements require them before the requested rudder angle is reached (i.e., in case a too long switching off is present). In the present case where the rudder rate is 5\u25e6/s, a sufficiently high value, the \u2018switching off\u2019 will not cause any appreciable effect. This effect is not considered in the present algorithm, but may need closer attention for low rudder rates (e.g., 1\u25e6/s) with fast ship response. Consider all possible ship configurations vis-\u00e0-vis the desired path at a time instant as shown in Fig. 3. The configurations may be classified as combinations of \u03b8e k > 0, \u03b8e k = 0 and \u03b8e k < 0 for the slope error and De k > 0, De k = 0 and De k < 0 for offset error, a total of 9 combinations. It may be kept in mind that \u03b4 > 0 makes the ship turn towards starboard side (towards right) and \u03b4 < 0 makes the ship turn towards port side (towards left). Therefore, for each combination, the sign of the command rudder angle \u03b4 can readily be figured out from Fig. 3 and these are summarized in Table 3. In order to develop the fuzzification and defuzzification processes, define a fuzzy membership function as \u03bc(X), where X = \u03b8e k or De k or \u03b4. There are many forms of the membership function and in the present work a triangular membership function (with values between 0 and 1) is adopted and this is sketched in Fig. 4. This function is given by \u03bc(X) = \u23a7\u23aa\u23a8 \u23aa\u23a9 0 for X XL, (X \u2212 XL)/(XC \u2212 XL) for XL < X XC , (XR \u2212 X)/(XR \u2212 XC ) for XC < X XR, 0 for X > XR. (10) The two error inputs may now be discretized into a variable number of fuzzy sets, described by linguistic variables, depending upon the accuracy required for control effectiveness"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003386_978-3-642-11615-5_12-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003386_978-3-642-11615-5_12-Figure2-1.png",
+        "caption": "Fig. 2. Biomechanical model of the orbit using strands and mechanical constraints",
+        "texts": [
+            " A linear system (called a Karush-Kuhn-Tucker system [1]) is solved at each time step to obtain the generalized velocities at the next step. ( M GT G 0 ) ( \u03a6(k+1) \u03bb ) = ( M\u03a6(k) + hf \u2212\u03bcg ) , (2) where \u03bb is the vector of Lagrange multipliers for the constraints. A direct method based on Gaussian Elimination is used to solve this matrix. Once the new velocities, \u03a6(k+1), are computed, the new positions are computed using Rodrigues\u2019 formula [10]. Each EOM is modeled as one or more strands. Other ocular structures are defined as mechanical constraints. The components in the model are illustrated in Figure 2 and described as follows. a. The globe is a spherical rigid body with a ball-and-socket joint allowing rotation in 3D. The translational movement of the globe is assumed negligible. Extraocular muscles exert torques on the globe through attachments. b. Each rectus muscle has two contractile elastic strands modeling the global and orbital layers. The EOM constitutive model will be discussed below. Two strands are chosen as this is sufficient to model the EOM path and mechanics along the principal axis that is functionally most important in moving the eye"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001141_j.jmatprotec.2008.02.071-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001141_j.jmatprotec.2008.02.071-Figure3-1.png",
+        "caption": "Fig. 3 \u2013 Profile of the novel-design rack cutter.",
+        "texts": [
+            " (34), the meshing equation becomes: = [ b0 \u2212 R cot ( 45 + R ) \u2212 HKW tan R + r1 1 ] (48) 2 Solving this meshing equation (48) and the locus equation (32) simultaneously, the generated tooth profile by cutting face V can thus be obtained. 852 j o u r n a l o f m a t e r i a l s p r o c e s s i n g t e c h n o l o g y 2 0 9 ( 2 0 0 9 ) 847\u2013855 Example 1. A spur-typed cutter has a circular pitch cp = 2.8 mm, with 12 flutes, an outside diameter of 11.391 mm, and a root diameter of 7.851 mm. The relevant parameters are shown in Table 1. The profile of the rack cutter is first designed as shown in Fig. 3. Using the mathematical model developed, the generated spur-typed cutter is shown in Fig. 4. Fig. 4 also shows that the developed mathematical model matches the simulation of the generated rack cutter. The result proves that the novel design of using the proposed rack cutter (hob cutter) is an effective and efficient way of manufacturing a spur-typed cutter. Fig. 5 shows the transverse section of the spur-typed cutter. This figure reveals that the same rack cutter with a different shift can produce different spur-typed cutters"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002006_j.cirp.2010.03.071-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002006_j.cirp.2010.03.071-Figure3-1.png",
+        "caption": "Fig. 3. Definition of the investigated angles in the experimental set-up.",
+        "texts": [
+            " The focus position of the adaptive mirror is set to zf = +21 mm as the built-up rate is better with the bigger spot size and thus the process is more efficient (see Fig. 2). However, also the smaller spot size (zf = 21 mm) gives good results, but at much lower built-up rates. The first experiments on variation of the surface angle are carried out via tilted workpieces. Two angles are defined that represent the tilting of the surface out of the horizontal plane (W) and the tilting of the nozzle out of the vertical position (N). As long as both values are equal, the nozzle is orientated perpendicular to the surface (see Fig. 3). The first row of experiments changes W = [08, 458] in steps of dW = 158 without changing N. At W = 458 the layer begins to show holes due to flowing of the molten layer material. The reason lies in angles with a big laser spot diameter (zf = +21 mm). the mass of the liquid metal. At certain angles (W) the gravitational force exceeds the surface tension. Certainly the wetting of unwanted portions of the surface is supported due to the nonperpendicular laser beam that now heats an elliptical shape"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001488_s12195-008-0027-5-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001488_s12195-008-0027-5-Figure2-1.png",
+        "caption": "FIGURE 2. Schematic showing the model variables and parameters. The model bacterium of radius R and side length Ls, is coated with a nucleation promoting factors (NPF) at a mean density concentrated to one pole. An individual filaments is assumed bound at the (+)-end to an NPF located at ri and anchored at position r0,i located at filament free-segment length L from the tip. The NPF has a spring constant, and the filament mechanics are determined by its bending stiffness. The position vector rp identifies the center of the rearward half-sphere, and the instantaneous orientation is identified by directors (orthogonal unit vectors) k1, k2, and k3. The NPF density depends on the contour position s, where s = 0 is taken as the boundary position between spherical and cylindrical regions of the pill-shaped bacterium.",
+        "texts": [
+            " A simulation of the propulsion trajectory therefore requires a model for force-dependent filament elongation, while accounting local variations in monomer concentration arising from diffusion limitations. The forces and torques generated by individual filaments depend on their orientations and positions on the microbial surface as well as how the filament stresses are transmitted into the highly cross-linked actin tail. To facilitate comparison to Listeria trajectories, the analysis is confined to a pill-shaped particle of radius R and side-length Ls (assumed equal to 2R) (Fig. 2). Let rp(t) identify the instantaneous position (identified by the center of the rear half-sphere) and the directors k1(t), k2(t), k3(t) (orthogonal unit vectors) identify the orientation. The net force and moments are assumed to be in balance with drag forces, such that b drp dt \u00bc F \u00f01\u00de br dki dt \u00bc T ki i \u00bc 1; 2; 3 \u00f02\u00de where b and br are the translational and rotational drag coefficients, respectively. As discussed below, the applied forces include those generated by individual filament (+)-ends (discussed below), as well as a steric force that prevents the microbial boundary from crossing the actin network",
+            " (Here and hereafter an \u2018\u2018NPF\u2019\u2019 refers to a functional filament nucleating and end-tracking unit at a surface density that corresponds to the local maximum density of filament ends. The actual surface density of NPF\u2019s molecules may exceed this value. For example, the end-tracking mechanism anticipates that at least two NPF molecules are required per filament tip11). To reflect the distribution of ActA on Listeria,35 the rearward half-sphere was assumed uniform with density q0, with the density on the sides decreasing exponentially with arc length s (see Fig. 2), i.e., q\u00f0s\u00de \u00bc q0e max\u00f0s;0\u00de=Ls with decay constant Ld. During random placement of NPF\u2019s (by random number generation), any new NPF placed within one filament diameter (7 nm) of an existing NPF was rejected and another was added elsewhere. In order to explore the possible effects an asymmetric distribution around the long axis, in some simulations NPF\u2019s were concentrated in the azimuth h according to the wrapped Cauchy density,14 q\u00f0h; s\u00de \u00bc qmax\u00f0s\u00de 1 e\u00f0 \u00de2 1\u00fe e2 2e cos h \u00f04\u00de where qmax\u00f0s\u00de \u00bc q0e max\u00f0s;0\u00de=Ls : As is presumably true for the actual pathogen, the polar concentration of NPF\u2019s facilitates symmetry breaking and the onset of directional motility in the simulations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001785_j.ijmecsci.2009.06.001-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001785_j.ijmecsci.2009.06.001-Figure10-1.png",
+        "caption": "Fig. 10. Model of an isolated cell wall. The end connections with its neighbours are replaced by vertical kv , horizontal kh and torsional springs ky .",
+        "texts": [
+            " Encouraged by the calculations on the upper bounds, a model of a pinned\u2013pinned isolated cell wall was considered because it represents the idealised end condition of a member in a truss. The resonant frequency is calculated as 5419 Hz, which is much less than the frequency of 8360 Hz where bending dominated modes appear in Fig. 8. This showed the need of a more realistic model. Since the ends of each cell wall possess three degrees-of-freedom in a plane (two translations and one rotation), a model for the individual cell wall is presented in Fig. 10. Note that this model is valid only for calculations of the cell wall resonance frequency. A vertical spring kv, a horizontal spring kh, and a rotational spring ky account for the stiffness at the ends. Since it is very difficult to have an estimate for these stiffnesses analytically, these stiffnesses were determined computationally by removing a cell wall from the overall cellular beam model and applying forces or moment appropriately. The values of kh, kv and ky were numerically calculated as 1:29 108 N=m, 1:16 106 N=m and 5"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000738_acc.2007.4283136-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000738_acc.2007.4283136-Figure1-1.png",
+        "caption": "Fig. 1. Formation configuration for two UGV.",
+        "texts": [
+            " Let F\u0304 d ij = [ldij \u03b7d ij ] T be the desired formation between robots i and j in a frame attached to robot i, where ldij \u2208 R+ is the desired distance and \u03b7d ij \u2208 [ \u03c0 2 , 3\u03c0 2 ] is the desired relative bearing angle. The actual formation for robot-pair i and j is described by F\u0304ij(t) = [lij \u03b7ij ] T , in which the relative distance is defined as lij(t) = \u2016ai \u2212 aj\u20162 , where \u2016\u00b7\u2016 2 denotes the standard Euclidean norm. The relative bearing is given by \u03b7ij(t) = \u03c0+\u03b6\u2212\u03b8i, where \u03b6(t) = arctan 2(yi\u2212yj , xi\u2212xj). Figure 1 shows the formation configuration for two UGV. The transformation between Fij and F\u0304ij is straightforward. The formation error for robot-pair i and j is then defined as eij = F\u0304 d ij \u2212 F\u0304ij . Different from conventional control which uses precomputed control laws, MPC is a technique in which the current control action is obtained by solving, at each sampling instance, a finite-horizon optimal control problem. Each optimization yields an open-loop optimal control sequence and the first control of this sequence is applied to the plant until the next sampling instance"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure1-1.png",
+        "caption": "Fig. 1 Normal sect",
+        "texts": [
+            " Additionally, this type of gearing exhibits a point contact rather than the line contact, which is usual in traditional involute helical gears, and the bearing width can easily be controlled by the curvature of the cycloidal curve. 2 Mathematical Models of the Circular Arc Cutting Edges In this model, two head cutters generate the corresponding sides of the gear tooth. Because each head cutter cuts only one side of the tooth flank, the cutter blades mounted on the same head cutter are either all inner or all outer blades that rotate with respect to the same rotation center. As shown in Fig. 1, the cutting edge of the MARCH 2010, Vol. 132 / 031008-110 by ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use c p a s r t S b b a H ion 0 Downloaded Fr utter blade on the front cutting plane can be designed with two arts: the circular edge along the curve M1M2 with a radius of rt nd a circular arc of the radius jk with its center at Cjk. The ubscript j may be replaced with i or o for inner or outer blade, espectively. The front cutting plane is normal to the direction of he relative cutting velocity, and the coordinate system a xa ,ya ,za is rigidly attached to the front cutting plane of the lade"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000725_acc.2008.4586578-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000725_acc.2008.4586578-Figure2-1.png",
+        "caption": "Fig. 2. Output Synchronization in SE(3)",
+        "texts": [
+            " If we now consider the velocity V b i as an input and the vector form of the rigid-body motion \u03a0i as an output, Lemma 1 says that the rigid-body motion in SE(3) (1) is passive from the input V b i to the output \u03a0i (This property is called a passivitylike property throughout this paper) in the sense defined in [16], since integrating (2) from 0 to T yields\u222b T 0 (V b i )T \u03a0idt = \u03c8(gi(T )) \u2212 \u03c8(gi(0)) \u2265 \u2212\u03c8(gi(0)). B. Output Synchronization in SE(3) Next we define output synchronization as follows. Definition 1 (Output Synchronization): A group of n rigid bodies is said to achieve output synchronization, if lim t\u2192\u221e\u03c8(g\u22121 i gj) = 0 \u2200i, j \u2208 {1, \u00b7 \u00b7 \u00b7 , n}. (2) By the definition of the function \u03c8, equation (2) implies the outputs of all the rigid bodies converge to a common value. From the definition of the output, it means that both of positions and orientations converge to a common value (Figure 2). In this paper, we present the velocity input V b i = [ e\u2212\u03be\u0302i\u03b8i 0 0 e\u2212\u03be\u0302i\u03b8i ][ vd e\u2212\u03be\u0302d\u03b8d\u03c9d ] +Ki \u2211 j\u2208Ni wij [ e\u2212\u03be\u0302i\u03b8i 0 0 I ][ pj \u2212 pi sk(e\u2212\u03be\u0302i\u03b8ie\u03be\u0302j\u03b8j )\u2228 ] , (3) where Ki = [ kpiI 0 0 keiI ] , kpi > 0 and kei > 0 are gains for the position error and attitude error respectively. In addition, \u03c9d is defined by \u03c9d := e\u2212\u03be\u0302d\u03b8d e\u0307\u03be\u0302d\u03b8d . vd and \u03c9d are desired linear and angular velocity. The present input (3) consists of two parts (feedforward and feedback laws). The first term is the feedforward part, which specifies a desirable behavior after the output synchronization is achieved"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003932_1077546312461026-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003932_1077546312461026-Figure3-1.png",
+        "caption": "Figure 3. Variation profile for i.",
+        "texts": [
+            " By setting _si \u00bc 0 for i \u00bc 1, 2, the equivalent control signals are determined to be u1eq \u00bc 1 b\u03021 _ud 2 1 _e1 2 1e1 f\u03021 n o \u00f013a\u00de u2eq \u00bc 1 b\u03022 \u20ac d 2 _e2 f\u03022 n o \u00f013b\u00de By satisfying the sliding conditions, the ki terms in equation (12) can be written as k1 1 1 \u00fe F1\u00f0 \u00de \u00fe 1 1 1 _e1 _ud \u00fe f\u03021 \u00f014a\u00de at NATIONAL CHUNG HSING UNIV on April 12, 2014jvc.sagepub.comDownloaded from k2 2 2 \u00fe F2\u00f0 \u00de \u00fe 2 1 2 _e2 \u20ac d \u00fe f\u03022 \u00f014b\u00de To prevent the controller from overreacting when the system is in the vicinity of the sliding surfaces, the values of the ki gains have been manipulated without violating the sliding conditions. This is done by linearly varying each i control parameter according to the distance, di, between the current location of the system and the ith sliding surface (see Figure 3). As a consequence, the i term is computed as follows i \u00bc Sat sij jffiffiffiffiffiffiffiffiffiffiffiffiffi 1\u00fe 2i p ! i \u00bc 1, 2 \u00f015\u00de where \u2018\u2018Sat\u2019\u2019 is a saturation function. A sliding mode observer is designed to accurately estimate the state variables of the surge and yaw motions of the ship. The available measurements are considered to be the heading angle along with theX andY coordinates of the ship with respect to the inertial reference frame. Both X and Y can be obtained from a global positioning system (GPS)while the yaw angle, , can bemeasured by an on-board gyro compass system (Fossen and Strand, 1999a)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003031_iros.2011.6094957-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003031_iros.2011.6094957-Figure2-1.png",
+        "caption": "Fig. 2. Reachable directions for a fingertip i: a) Workspace \u03c6i for the fingertip; b) Directions of forces that the fingertip can apply on the object.",
+        "texts": [
+            " For each pose, a collision detection between the object and the hand is performed, with the fingers in the configuration of maximum aperture of the hand. If there is no collision, then the intersection \u03c8i between the object and the workspace \u03c6i for each finger is computed. The set \u03c8i includes all the points on the object reachable for each finger i. The next step creates the sets \u03c8\u2032 i of intersected points whose normals are within the potential directions of force that the fingertip can apply, as shown in Fig. 2. If at least two sets \u03c8\u2032 i are not empty, then it is verified whether the reachable points lead to a force closure grasp. The steps in the algorithm are: Algorithm 1: Computation of the graspability map 1) Voxelize the parallelepiped delimiting the possible locations of the hand base frame around the object 2) Define a set \u0393 of potential locations and orientations for the hand base frame 3) For each pose of the hand base frame in \u0393 a) Check for collisions between the hand and the object. If there is a collision, discard the pose b) For each finger i compute \u03c8i = \u03c6i \u2229 \u2126 c) Obtain the sets \u03c8\u2032 i \u2282 \u03c8i of points with normals within the directions of force that each fingertip can apply d) If at least two sets \u03c8\u2032 i are not empty Verify the force closure condition Else Discard the pose 4) Return all the poses in \u0393 that lead to FC grasps Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002878_iccis.2010.5518579-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002878_iccis.2010.5518579-Figure4-1.png",
+        "caption": "Fig. 4. Path following geometry",
+        "texts": [],
+        "surrounding_texts": [
+            "The performance of the designed control system must be evaluated on the nonlinear counterpart and finally by flight test. The schematic is shown as below: we find that the performance still satisfies our requirements around the trimming point although it presents decrease in some extent when using the nonlinear model in simulation and even further in flight test. It is due to the accuracy of our nonlinear dynamic model."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003727_j.engfailanal.2011.06.004-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003727_j.engfailanal.2011.06.004-Figure2-1.png",
+        "caption": "Fig. 2. Geometry modeling of the cooling fan including fan blades, hub plate and U-bolts.",
+        "texts": [
+            " The etching solution was a . All rights reserved. 9x5684; fax: +98 611 3336642. n). mixture of 60 ml HCl and 20 ml HNO3 (Aquaregia solution). The fracture surface of the U-bolt was investigated by visual inspection and scanning electron microscopy (SEM), using secondary electron detectors. A structural analysis of U-bolts of a cooling tower fan is performed using finite element modeling. SolidWorks software is used for geometry modeling of the cooling fan including fan blade, hub plate and U-bolts as shown in Fig. 2. Finite element software ANSYS 12 is used for stress analysis. The elements used in U-bolt model are all 8-node solid brick elements Solid45. Finite element model for stress analysis is consisted of 10,000 elements and 16,200 nodes for each U-bolt as shown in Fig. 3. All parts of the cooling fan assembly except the hub plate are considered flexible. Since blade model must transfer forces to the U-bolts, it is not considered rigid in this analysis. To model the contact between the U-bolts, blade and hub surfaces, CONTA174 and TARGE170 elements are used"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001741_17543371jset75-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001741_17543371jset75-Figure2-1.png",
+        "caption": "Fig. 2 The \u2018grip simulator\u2019 used to generate a level of restrictive torque about the racket handle",
+        "texts": [
+            " The inbound velocity of the ball in the y direction was varied by rotating the racket mount such that incoming balls were obliquely incident to the racket face. The magnitudes of the ball\u2019s velocity in the x and y directions could therefore be altered by changing the ejection speed Proc. IMechE Vol. 224 Part P: J. Sports Engineering and Technology JSET75 at UNIV OF WISCONSIN-MADISON on March 25, 2015pip.sagepub.comDownloaded from of the BOLA machine and the rotation of the racket mount. The racket mount consisted of a handle clamp, limiting clutch, and universal joint (Fig. 2); the racket was attached by the handle and allowed to rotate freely about two axes owing to a universal joint attached to the top of the arrangement. The universal joint was added to relieve stresses in the racket and to prevent breakages from impact. It was assumed from prior research that extra handle mass and restricted racket movement due to the mount would not affect the postimpact ball velocity [1, 8, 9]. The post-impact racket movement was not measured during the experiment. The limiting clutch, manufactured by Cross+Morse [10], provided a level of restrictive torque"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001297_ems.2008.85-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001297_ems.2008.85-Figure5-1.png",
+        "caption": "Figure 5. Bearing dimension and characteristic defect frequencies.",
+        "texts": [
+            " Td is the fault characteristic period, i.e., the reciprocal of the fault characteristic frequency fd, which can be calculated from the geometry of the bearing and its rotational speed. The model has a number of userspecified parameters [12]. Local defects or wear defects cause periodic impulses in vibration signal. Amplitude and period of these impulses are determined by shaft rotational speed, fault location, and bearing dimensions. The frequency of these impulses, considering different fault locations as in Figure 5. Fundamental cage frequency is given by [10]: ))cos(1( 2 \u03b1 D df f s c \u2212= (3) Ball defect frequency is two times the ball spin frequency and can be calculated as: ))(cos1( 2 2 2 \u03b1 D d f d D f sdb \u2212= (4) Inner race defect frequency is given by: ))cos(1( 2 )( \u03b1 D dnf ffnf s esid +=\u2212= (5) Outer race defect frequency is given by: ))cos(1( 2 \u03b1 D dnf nff s eod \u2212== (6) where fs is the shaft rotation frequency in rads-1, n is the number of rolling elements, d is the roller diameter, \u03b1 is the contact angle and D is the pitch diameter of the bearing"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000360_cima.2005.1662303-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000360_cima.2005.1662303-Figure1-1.png",
+        "caption": "Fig. 1. Diagram of the DC motor system constant, 2 E R , R+ is the space of positive real constants. The sliding surface given in (8), however, may cause steady-",
+        "texts": [
+            " Sliding mode The motors have been extensively used in several industrialcontrol (SMC) method, first proposed in the early 1950s, has applications [21,22,23]. There have been considerable been demonstrated to provide an effective control action to ddevelopments in nonlinear control schemes for the dc motorschallenge system uncertainties and disturbances with good [21,24,25]. This has attracted extensive researches in the fieldrobustness and sufficient performance [3,4]. of control engineerin A key point in the design of sliding mode controllers is to The g gTerough diagram of the system studied iS shown in Fig. 1.introduce a proper sliding surface so that tracking errors and The electrical and mechanical equations representing present output deviations can be reduced to a satisfactory level. dc motor system can be given as [26]: Unfortunately, an ideal sliding mode controller has a discontinuous switching function and it is assumed that the d control signal can be switched from one value to another va(t Laia +Raia +KmWm (1)dt infinitely fast [1,5]. In practical systems, it is impossible to achieve infinitely fast switching control because of fmite time Kmia(t) = Jm-ia + Rm\u00b0m (2) delays for the control computation and limitations of physical dt actuators [6,7]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000469_14644193jmbd78-Figure14-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000469_14644193jmbd78-Figure14-1.png",
+        "caption": "Fig. 14 Forces acting on the piston during successive strokes",
+        "texts": [
+            " At the mid span of piston motion, the sliding velocity is higher. This tends to pull more lubricant into JMBD78 \u00a9 IMechE 2007 Proc. IMechE Vol. 221 Part K: J. Multi-body Dynamics at NATIONAL UNIV SINGAPORE on June 28, 2015pik.sagepub.comDownloaded from the contact by entraining action. Thus, a larger film thickness is observed in each of the cycles at mid span, except at the power stroke. During the power stroke, the piston is pushed towards the thrust side by the tangential force component (Fs in Fig. 14) in the connecting rod. This results in a smaller film thickness. As the film thickness is very small during the power stroke, the viscous friction force is much higher. During other strokes the film thickness is comparatively larger and the friction force is correspondingly lower. The maximum combustion gas force is applied soon after the TDC. At this point, the side force (thrust force) attains its maximum value. Furthermore, near to the TDC the sliding velocity is low, leading to a smaller film thickness soon after the TDC as shown in the figure",
+            " At the end of the exhaust stroke, the film thickness is at its maximum and the friction force reaches its minimum value. The rapid oscillatory variations in force shown in this figure and those in Fig. 13 have no physical significance. These occur in the vicinity of reversal positions and are due to mathematical inaccuracy due to integration time step size, which should ideally be reduced. However, such a reduction leads to significant rise in computation times. It is interesting to examine force variation in the piston during its motion. In Fig. 14 the forces due to combustion gas force, Fp and the inertia force, Fi, acting on the piston during its downward motion during the suction stroke are shown. These two forces have to be balanced by the force, FR, acting on the connecting rod. Even though vertical component, Fd on the connecting rod is balanced by Fp and Fi, the horizontal component, Fs has to be balanced by the thrust reaction from the cylinder wall. Viscous friction force acting at the piston\u2013wall interface is given as FF . The horizontal force variation in the piston is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002000_msf.618-619.159-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002000_msf.618-619.159-Figure2-1.png",
+        "caption": "Figure 2. Schematic of the LAM set up.",
+        "texts": [
+            " Argon gas was introduced to the cutting zone to prevent oxidation of the laser irradiated surface and to prevent contamination of the laser optics. The laser irradiates a circular spot on the workpiece, upstream of the cutting tool in the radial direction and upstream of the tool in the axial direction. These lead distances can be changed and allow the workpiece to be preheated before it is cut. The lead distance is chosen so that both the temperature gradient across the cutting tool and the temperature of the resulting workpiece are minimized. A schematic of the LAM experimental setup is provided (Figure 2). Experiments have shown a 60% improvement in material removal by heating the material at an optimized temperature (Figure 3) as predicted by modeling. Empirical results were obtained without cutting fluid at the same tool life. Largely, this means that 60% more material can be removed at the same tool life without the environmental issues involved with discarded cutting fluid. Results showed that no improvements to removal rate (Figure 4) could be made with the addition of temperature. To overcome this limitation a hybrid system was devised where the cutting tool was actively cooled simultaneously"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002678_j.wear.2011.10.005-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002678_j.wear.2011.10.005-Figure2-1.png",
+        "caption": "Fig. 2. RCF test elements: (a) Rex 20 and M50 balls, races and rods and (b) M50 balls against a Si3N4 rod with M50 outer races inside the vacuum chamber.",
+        "texts": [
+            " The drive motor is mounted outside the chamber and motor torque is delivered to the rod inside the vacuum chamber using a ferro-fluidic rotary feed-through device similar to Kurt J. Lesker VacuumTM part number FE121099. The races are press fit into the test fixture and held stationary while the balls rotate between the fixed races and the rotating rod. High vacuum is applied using a VarianTM V-81M turbo pumping system as shown in Fig. 1b. The test rig used for all RCF tests is a modification of the threeball\u2013rod test of Hoo [11]. The test head and related components were fabricated from 304L stainless steel and positioned inside a UHV chamber as shown in Fig. 2. A carrier cage was not used for these tests. Instead, 5 balls are used without the carrier cage for testing with 12.7 mm diameter balls as described in configuration 1 of Table 1. Six 7.94 mm diameter balls were tested as well and are presented in configuration 2 of Table 1. Si3N4 rods were also tested as presented in configuration 3 of Table 1. The Si3N4 rods were purchased from CeradyneTM Inc., and were processed using the EKasinTM method from Ceradyne. The EKasin process involves gas pressure sintering and the materials from this process were chosen for their high temperature and high thermal shock resistance capability",
+            " 1, the system was outgassed for 12 h before starting the test. All tests were carried out in UHV conditions in the ranges of 10\u22124\u201310\u22125 Pa. Through previous testing using the RCF rig in Fig. 1, the onset of a spall on at least one ball is detectable over a vibration range of 0.22\u20130.35 g with the accelerometer mounted as shown in Fig. 1. The failure threshold for all tests was the same, 0.35 g, as tles to a range of 0.06\u20130.15 g until the failure threshold is exceeded. The temperature of the test fixture shown in Fig. 2b increases with Table 1 Ball, rod, and race configurations for RCF high vacuum testing. Config Ball material Rod material Race material Lubrication Ball size Number of balls Number of tests 1 M50 steel Rex 20 M50 steel Silver 12.7 5 33 2 ANSI T5 Rex 20 Rex 20 Silver 7.94 6 27 3 M50 steel Si3N4 M50 steel Silver 12.7 5 17 4 M50 steel Rex 20 M50 steel Silver 12.7 3 2 Table 2 Hardness and material property data for ball, rod, and race components. Material property 9.53 mm diameter Rex 20 rod 9.53 mm diameter Si3N4 rod 7.94 mm steel ANSI T5 ball 12.7 mm steel M50 ball Rex 20 race M50 race v t t r t t i e a c T fi T S HRC (measured) 62.9 74.6 Elastic modulus, GPa 235 310 Poisson ratio 0.29 0.25 ibration and tended to increase over the length of the test. As he silver (lubrication) is depleted from the ball surface, the sysem vibration increases as well. The peak temperature at failure is epeatable at about 120 \u25e6C, as measured by a thermocouple in conact with the top race in the location shown in Fig. 2b. For most of he test, the temperature ranges from 86 \u25e6C to 100 \u25e6C as measured n the location shown in Fig. 2b. The ball temperature at failure was stimated to be approximately 160 \u25e6C using an optical pyrometer nd for 12.7 mm M50 balls. Test information data such as ball\u2013rod ombinations, hardness, and surface roughness are presented in ables 1\u20133. Figs. 3\u20135 contain cycle versus contact stress for three test congurations presented in Table 1. These curves were calculated in able 3 urface roughness Ra data in microns for ball, rod, and race components. 12.7 mm M50 ball M50 race 7.94 mm T5 ball 0.32 0.37 0.05 61",
+            " As soon as the input to the third-body volumes ceases, the third-body volume coverage X(t) diminishes resulting in asperity-to-asperity contact such that friction and vibration increase and the stopping vibration threshold of 0.35 g for the RCF test is exceeded. Table 5 presents the wear coefficients assumed in Eqs. (5) and (6). These values of wear coefficients are within the range and order of magnitude of those tested between bearing steels like Rex 20 and silver in UHV [12] and those tested between Si3N4 and silver [13], and are used here for trending only. Table 6 contains normal load and contact area calculations from the RCF test setup of Fig. 2. Eq. (5) is plotted in Fig. 10 using the material properties, wear coefficients, and loads presented in Tables 2, 5 and 6. Post-test thickness measurements of the silver remaining on the balls suggests a third-body coverage steady state value, Xss, Table 5 Wear coefficients used in Eqs. (5) and (6). K (m2/N) Kbc Kbr KbEc KbEr Test configuration 3: Si3N4 rod, M50 ball and race 1.0E\u221215 2.0E\u221216 1.0E\u221215 5.0E\u221217 Test configuration 1: Rex 20 rod, M50 ball and race 1.0E\u221215 2.0E\u221215 1.0E\u221215 2.0E\u221217 Table 6 Normal force and third-body contact area calculations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003078_rspa.2010.0449-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003078_rspa.2010.0449-Figure1-1.png",
+        "caption": "Figure 1. Schematic of a rod undergoing deformation from its straight state reference configuration, inset shows forces and moments acting at the end cross sections of a rod segment.",
+        "texts": [
+            " Under this condition, a fibre becomes more stable as it is sheared, hence it never buckles. The later simulates a fibre in a displacement controlled test in which a compressive preload is applied in the beginning. For this case, we show that the normal force on a fibre under initial compression can switch to tension, consistent with the observations of K.A.S. (a) Geometry and governing equations The undeformed rod is assumed to be straight and the arc-length coordinate of a material point on the centre line of the undeformed rod is specified by S (figure 1). The position vector of the material point S after deformation is denoted as R(S). A local coordinate system with orthonormal basis vectors {a\u0302 i} is attached at each cross section, with a\u03023 aligned with the centre line tangent, while the other two vectors are aligned with the principal flexural axes of a rod\u2019s cross section. This body-fixed frame describes the orientation of a cross section with respect to the inertial frame {e\u0302i}. The curvature vector K is the rotation of Proc. R. Soc. A (2011) the body-fixed frame with respect to the inertial frame per unit of undeformed arc-length S , ( va\u0302 i vS ) {e\u0302i} = K \u00d7 a\u0302 i "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002348_1350650111427508-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002348_1350650111427508-Figure5-1.png",
+        "caption": "Figure 5. The impact hammer mounting.",
+        "texts": [
+            " In order to avoid the dynamic misalignment problem exhibited during previous tests,14 the housing is mounted on three flexible stingers fixed on a base plate. The stingers do not interfere with measurements because their stiffnesses are one order of magnitude lower than those of the tested bearing. The static load is vertically applied on the bearing by a manually driven bolt and nut system and spring. Its value is measured with a dynamometer. The impact hammer is attached to a platform fixed on the milling table and has a rotational degree of freedom around a pivot. This simple mounting enables to apply a repeatable 45 inclined impulse load (Figure 5). Impacts are hence delivered in two orthogonal directions that coincide with (X,Y) in Figure 2. The altitude of the hammer launching and its head (shape and material) will control the energy transferred by the impact. Inductance proximity probes are used for measuring the relative displacements between the rotor and the bearing. They are mounted on the right and left axial planes of the housing (two in each plane). The impacts are collinear with the proximity probes. The housing is also equipped with four accelerometers (two in each end plane)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001141_j.jmatprotec.2008.02.071-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001141_j.jmatprotec.2008.02.071-Figure7-1.png",
+        "caption": "Fig. 7 \u2013 Definitions of the cutting angles.",
+        "texts": [
+            " (29) and 42), respectively, are the same as the x- and y-components f the equation depicting the undercut portion shown in Eqs. 31) and (46). xample 2. In Example 1, when the shifted distance c = 0.0, he radius of the intersection point is rP = 5.725, which is larger han the radius of the outside diameter, r = 5.696. This confirms hat the design of the rack cutter has satisfied the requirement f a spur-typed cutter. .2. Determination of the cutting angles he cutting angles of the end section profile shown in Fig. 7 ffect the cutting performance significantly. In this section, he cutting angles of a spur-typed cutter manufactured by the ovel hob cutter are investigated. Point A is the intersection oint of the undercut portion (region IV) and the curve on the op land of the spur-typed cutter, point B is the intersection oint of the chamfered angle edge and the top land curve, nd point E is the intersection point of the left cutting edge nd the chamfered angle cutting edge. T\u2032 A and Ta are the tanential and positional vectors of point A in the S1 coordinate ystem, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003017_s13369-010-0022-8-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003017_s13369-010-0022-8-Figure9-1.png",
+        "caption": "Fig. 9 The pole placement region",
+        "texts": [
+            "1 Pole-Placement Region The required pole-placement region can be designed using half-plane, disk, and conic sector regions characterized by S = (\u03b1, r, \u03b8). A minimum delay rate \u03b1, a minimum damping ratio \u03be = cos \u03b8, and a maximum undamped natural frequency \u03c9 = r sin \u03b8 is ensured when such a region is proposed. The values used for the region are as follows: the left haft-plane \u03b1 = \u221210, the disk center is at zero and the radius r = 200, and the conic sector origin is at zero, and the half inner angle \u03b8 = 3\u03c0/2, as shown in Fig. 9. The poles are chosen such that the settling time is less than one second. The rationale for this settling time specification comes from the observed quasi-frequency of ocean waves (taken as disturbance) as it can be seen in the observed experimental data shown in Fig. 17. A wide variation in this disturbance is expected with seasonal and geographic changes and is also due to the tracking orientation of the asymmetric payload. A safety factor of ten is assumed for this purpose. For the mentioned rise time, the desired natural frequency comes out to be 2 rad/s"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure5-1.png",
+        "caption": "FIGURE 5. The wheel and axle (Mechanics 2.1, 2.10). On the left is Drachmann\u2019s drawing (The Mechanical Technology, p. 51) made from Ms L; on the right is the figure from Heronis Alexandrini opera, vol. II, p. 96.",
+        "texts": [
+            " arrikat) applied at its end (B) is identified with the force applied to the circumference of the larger circle, the weight (t iql) to be lifted (G\u030c) with the weight on the arm of the smaller circle, and the fulcrum (D) with the centre of the two circles. Since the ends of the lever trace out arcs on concentric circles as the weight is lifted, the analysis of 2.7 can be applied directly: if the ratio of the length of the longer arm BD to that of the shorter arm DA is equal to the ratio of the weight G\u030c to the moving force applied at B, the lever is in equilibrium; if it is greater than the ratio of G\u030c to B, the force will lift the weight.33 The reduction of the wheel and axle to the concentric circles is just as direct (2.10; Fig. 5). The wheel corresponds to the larger circle and the axle to the smaller; the weight (t iql) is hung from the axle and the moving force (al-quwwat al-muh. arrikat ) applied at the circumference of the wheel.34 32 Dioptra, ch. 37 (Opera, vol. III, p. 310.26\u20137); see below, pp. 34\u201335. Pappus, Pappi Alexandrini collectionis quae supersunt, VIII, vol. III, p. 1066, makes the same point in the same language. 33 On the second use of the lever, discussed in Mechanics 2.9, see below, pp. 43\u201345. 34 Heron himself remarks in 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure7.3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure7.3-1.png",
+        "caption": "Fig. 7.3-1 Laplace domain representation of a single-degree-of-freedom system, where m= 10, c = 1 5. , k = 3 : (A) real part, (B) imaginary part [7.2]. (This graph is reproduced with permission of McGraw-Hill Companies.)",
+        "texts": [
+            " Similarly, with zero initial conditions, the ratio of the transform of the acceleration to that of the forcing function, denoted by Ha(s), is H s s H s s s cs ka d( ) ( )= = + + 2 2 2m (7.3-6) Ha(s) is alternately called the acceleration transfer function, or the accelerance, or the inertance. Since transfer functions are complex-valued functions of the complex independent variable s, they have real and imaginary parts. Th us, a transfer function can be viewed as a function of two variables and represented as a surface. Fig. 7.3-1 shows a typical graphical representation of the receptance transfer function of a single-degree-of-freedom system [7.2], where the mass, damping, and stiff ness values are m =10, c =15. , and k = 3; and where and are, respectively, the real part and imaginary part of the Laplace variable s. Th at is, s j= + . Th e Matlab code for exhibiting the surface is clear all m=10; c=1.5; k=3; a=-c/(2*m); b=sqrt(k/m-c^2/(4*m^2)); rot=60; lambda(1)=a+j*b; lambda(2)=a-j*b; [sigma,omega]=meshgrid(-0.2:0.01:-0.05,-1:0.02:1); s=sigma+j*omega; H=(1.0/10).*(1.0./((s-lambda(1)).*(s-lambda(2)))); view=[rot,30]; VibrationAnalysis_txt.indb 221 11/24/10 11:47:40 AM Principles of Vibration Analysis 222 mesh(real(H),view); pause mesh(imag(H),view) Th e defi nitions of undamped natural frequency, damped natural frequency, and the damping ratio are all relative to information represented by Fig. 7.3-1 [7.2]. Th e projection of the pole locations onto the plane of zero amplitude yields the damping coeffi cient, damped natural frequency, undamped natural frequency, and the damping ratio. Recall that, once the poles are known, the undamped natural frequency and damping ratio can be calculated from Eqs. 6.2-25 and 6.2-26. Th at is, n r r= ( ) +( )Re( ) Im( ) 2 2 (7.3-7) VibrationAnalysis_txt.indb 222 11/24/10 11:47:41 AM Transfer Functions and Frequency Response Functions | Chapter 7 223 and = \u2212 Re( )r n (7"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000782_tmag.2008.2002607-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000782_tmag.2008.2002607-Figure2-1.png",
+        "caption": "Fig. 2. Analysis model of PMSM/G. (a) PM rotor with diametrical magnetization and (b) a three-phase winding stator with a slotless iron core.",
+        "texts": [
+            " Consequently, the operating torque according to the driving mode is defined as: 1) motoring mode for rotating the flywheel ; 2) idling mode without the generating loads ; 3) generating mode with the electrical loads . Using Laplace\u2019s transformation of (1), the rotational speed in the frequency domain is given as (2) where is the initial angular speed of the flywheel rotor. In an effort to calculate the operating torque, this paper presents the analysis model of a manufactured PMSM/G with a PM rotor mounted on the flywheel and slotless three-phase winding stator as shown in Fig. 2. For a one-motor/generator structure, the PM rotor with the diametrically magnetized array and the slotless stator with the double-connection winding are idealized as a 2-D polar coordinate system in distribution of the -direction. Here, the radii of each material are given by the normal component from the shaft origin. This analysis model is assumed to have a depth of so that the end effects in the -direction are ignored. Letters , II, III, and IV indicate material regions, and its magnetic properties are applied from the permeability defined as in the air region and in the iron region",
+            " Generally, the magnetic field is obtained from Maxell\u2019s equation in relation to the magnetic-field intensity , magnetic flux density , and current density . Here, the magnetic flux density can be represented with the magnetic vector potential as since the divergence of the curl of the magnetic vector potential is identically zero, Ampere\u2019s law, expressed as , is rewritten from the magnetic vector potential inside each material, that is, [1] and [2] inside air and iron inside magnet (3) where H/m is the permeability of free space. In Fig. 2(a), the PM rotor with the two pole can be predicted so that the diametrical magnetization has radial and -components. This magnetization is expressed as Fourier series with only the fundamental component as follows: (4.1) and and (4.2) where represents the angular position of the PM rotor in the fixed stator frame and the magnetization amplitude of PM is given by from the remanence . In the two-dimensional cases where the fields lie, the magnetic vector potential is purely -directed in distribution of the -direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000923_bfb0110378-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000923_bfb0110378-Figure3-1.png",
+        "caption": "Fig. 3. A rigid body controlled by two body fixed forces.",
+        "texts": [
+            " It admzts q as a fiat output even when ~ zs singular-: indeed, u can be expressed in function of q, (t by the computed torque formula ( d (O~O) OL D(q , ( t ) ) . u = M(q) -1 ~ q If q is constrained by c(q) = 0 the system remains fiat, and the flat output corresponds to the configuration point in c(q) = O. Example 9 (P lanar rigid b o d y wi th forces) Consider a planar rigid body moving in a vertical plane under the influence of gravity and controlled by two forces having lines of action that are fixed with respect to the body and intersect at a single point (see figure 3) (see [78]). Let (x,y) represent the horizontal and vertical coordinates of center of mass G of the body with respect to a stationary frame, and let 0 be the counterclockwise orientation of a body fixed line through the center of mass. Take m as the mass of the body and J as the moment of inertia. Let g ,~ 9.8 mfsec 2 represent the acceleration due to gravity. Without loss of generality, we will assume that the lines of action /or F1 and F2 intersect the y axis of the rigid body and that F1 and F2 are perpendicular"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001579_1.5061435-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001579_1.5061435-Figure10-1.png",
+        "caption": "Fig. 10 Animated graphics of an inside view during the hardening process within the turning machine",
+        "texts": [
+            " Before the final machining steps the heat treatment is done with the part being in the opposed spindle. Because of laser and process safety the laser process occurs in a separate chamber what moves in on the dividing wall to the engine cabinet (Fig. 8). The spindle with the part moves into the chamber leaving a circumferential gap of about 0.5 mm between spindle and chamber inlet at the faceplate. Opposite of the part the laser beam is delivered under an angle of about 45\u00b0 to the turning axis within a hermetic telescoped tube (Fig. 9, Fig. 10). That tube and the process chamber are filled with a continuous flow of cleaned air to keep out moisture, dust and particles. That enables a parallel machining at the main spindle while the hardening occurs. Laser optics is protected effectively. In regular manufacturing conditions the cover slide needs a cleaning just weekly. The indication of a necessary cleaning comes from the temperature control unit \u00bbLompocPro\u00ab [7]. If the laser power progress of a process reaches the upper limit characteristics a message is given to the operator",
+            " Communication between laser, machines and Page 97 of 167 Poster Presentation GalleryICALEO\u00ae 2008 Congress Proceedings central controller is done via bus system. The beam switch controller enables priority driven choice of the beam path. Every kind of the parts is stored with its own recipe of machining time, handling time and laser time. The usage of the laser can be optimized and the queue time can be minimized. At the first step two turning machines and one shielding gas hardening machine are connected to the laser. A third turning machine will complete the line finally to get maximum output. Fig. 10 shows the setup of the laser hardening system. A great variety of valve cones and valve seats are manufactured in that production line finally since the mobile hydraulic valve comes in different sizes and with different options. The principle geometries are similar but there is a variation in size and position. At the turning machines the different CNC programs have their specific laser subroutines. At the shielding gas hardening machine for the valve seats different workpiece holders are used. The program routine is the same for all valve seats"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure4-1.png",
+        "caption": "Fig. 4. The part of wing FEM model",
+        "texts": [
+            " Thanks to that, there is the correct stress distribution on bulkheads near rib no 21. Methods for FEM analysis of riveted joints of thin walled aircraft structures 945 The wing is a torsion box construction. The wing torsion is made from spars, ribs and skin panels stiffened with stringers. The spars consist of upper and lower flanges of T-section and webs reinforced with struts. Ribs consist of upper and lower flanges, webs reinforced with struts and connectors. The part of wing FEM model is shown in figure 4. Part of upper skin was hidden to show internal structure. The model was prepared in MSC PATRAN software, calculation were made with MSC NASTRAN software. The model was used to gain boundary conditions for more accurate models. On a basis of previous analyses [12,13] rivets weren\u2019t represented. Shell elements (Quad4, Tria3) and linear material models were used. Jerzi Kaniowski et al. 946 The boundary conditions were taken on a basis of the operational data. All moments were converted into forces and applied to the structure"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003788_secon.2012.6196934-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003788_secon.2012.6196934-Figure2-1.png",
+        "caption": "Fig 2. X-configuration and + - configuration",
+        "texts": [
+            " In order for the quadrotor to move about the roll axis, the throttles of the other side rotors (right or left) are increased, while reducing the same side rotor throttles. For movement about the pitch axis, the front or back rotors are increased or reduced in the same way as for roll. In case of movement about the yaw axis, the counter-clockwise rotating rotors throttles are increased for rotation of the vehicle in counterclockwise direction and the same holds good for clockwise rotation as well. The vehicle has two different configuration in which it can be flown, the \u2018X\u2019 configuration and the \u2018+\u2019 configuration. Figure 2 shows both the configurations. The experiments described in this paper were based on an X-configuration quadrotor as they are more stable compared to the + configuration, which is a more acrobatic configuration. 978-1-4673-1375-9/12/$31.00 \u00a92012 IEEE There has been work done on quadrotors in areas like control systems [1 \u2013 9], which apply different control algorithms on a quadrotor in order to stabilize it. There have also been journals which describe the complete control mechanism and the different sensors used [8]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000001_1.2032993-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000001_1.2032993-Figure2-1.png",
+        "caption": "Fig. 2 Wedge gap between sun roller and outer ring",
+        "texts": [
+            " The outer ring contains an inner cylindrical raceway and is connected through a spoke web to an output shaft that is supported by a double row bearing on the carrier housing. The sun roller contains an outer cylindrical raceway and is integrated with an input shaft that is supported by a ball bearing on the carrier. The sun roller is set eccentric to the outer ring by amount e, known as the eccentricity. Thus a wedge gap is formed in the annular space between the outer ring and the sun roller. For the orientation shown in Fig. 1, the wedge gap assumes a large end at top and a small end at bottom, as illustrated in Fig. 2. The loading planet is assembled in the wedge at the top position with its outer surface in frictional contact with the inner cylindrical raceway of the outer ring and the outer cylindrical raceway of the sun roller. The loading planet is \u201celastically\u201d supported through a bearing on a pin shaft that is fixed to the carrier. An elastic insert or a clock spring is provided between the pin shaft 858 / Vol. 127, OCTOBER 2005 rom: http://tribology.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/ and the inner race ring of the support bearing, allowing the loading planet to move from its assembly position should any tangential pulling force exerts on the planet",
+            " The following equations relate the maximum contact stress to the material and lubricant properties, and geometrical size and configuration of the friction drive: ph = 1 2Rs E Ms ofcls 8 where = 2 + 2 + + 2 + 1 cos 1 \u2212 cos 2 + 2 + 2 cos \u2212 1 , = e Rs Rs is the radius of the sun roller raceway, ls is the width of the raceway. Ms is torque applied to the sun shaft, E is the effective Young\u2019s modulus, e is the eccentricity, the planetary ratio defined as the diameter ratio of the outer ring raceway to sun roller raceway. is the azimuth position of the support planet see Fig. 2 . The allowable stress level depends on the service life expectancy of the friction drive and, most importantly, on material strength. Modern steel making process has significantly improved impurity of steels. As a result bearing grade steels routinely work under high contact stress up to 4 GPa 580 ksi . However, to ensure a long endurance life of a friction drive over a wide range of service conditions, it is recommended that the nominal operating contact stress be under 2 GPa 300 ksi . Finite Element Modeling"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002341_j.optlastec.2011.04.010-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002341_j.optlastec.2011.04.010-Figure3-1.png",
+        "caption": "Fig. 3. 3-D finite elements mesh generation of workpiece.",
+        "texts": [
+            " The 3-D numerical models are developed using the commercial FEM code ANSYS to predict the temperature and stress distributions in workpiece, specifically in the RMZ. Numerical simulation is decoupled to two steps, the temperature field result came from heat transfer analysis in the first step is used as input to the mechanical analysis, and the thermal solid elements with eight nodes Solid70 are converted to Solid45 with eight nodes for the mechanical analysis. The thermo-physics material properties are considered as dependent on temperature. As shown in Fig. 3, for the appropriate result of compromise between computing time and accuracy, a dense mesh is used around the fusion line and a coarser mesh is adopted for the rest. A 3-D Cartesian coordinate system is established on workpiece with x-axis along the longitudinal (melting) direction, y-axis along transverse direction and z-axis along the thickness direction. The governing partial differential equation of LSM process can be written as [26] @2T\u00f0x,y,z\u00de @x2 \u00fe @2T\u00f0x,y,z\u00de @y2 \u00fe @2T\u00f0x,y,z\u00de @z2 \u00bc 1 a\u00f0T\u00de @T\u00f0x,y,z\u00de @t \u00f01\u00de where let a(T)\u00bck(T)/r(T) c(T), which is called the thermal diffusivity; k(T), r(T) and c(T) represent the thermal conductivity, density and specific heat of the material, respectively, which depend on the temperature T"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003103_sis.2011.5952567-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003103_sis.2011.5952567-Figure1-1.png",
+        "caption": "Fig. 1 The gray circle is the body of the robot in overhead view, on which the short rectangles represent the IR sensors. There are 12 IR sensors evenly arranged around the robot. D is the max sensing distance of IR. The outside dashed circle stands for the range of wireless communication that is approximated as a standard circle, and the max range, R, is limited not more than D by lowering the power supply.",
+        "texts": [
+            " Because the small aggregate has shorter lifetime than the large aggregate, robots in small aggregate are easy to become free moving and can participate in large aggregate. Once all the robots aggregate together, the timers of robots are disabled and the whole process of aggregation is over. Moreover, our algorithm is described in detail in the rest of this section from the view of an individual robot. In our algorithm, all the robots are identical. The robot is mobile with limited ability of interaction including IR sensing for detecting objects and wireless communication for communicating with other robots. The model of robot is illustrated in Fig. 1, where the most important component is the model of interaction. Total 12 IR sensors are evenly placed around the robot with a max detecting distance D. The range of wireless communication is approximated as a standard circle, where the max range is R that is calculated as the distance from the outside of robot to the boundary of the circle. Since the range of wireless communication is limited to be short by lowering the power supply, both the IR sensing and the wireless communication are in short range, where R is not more than D, so that the \u201cRequest-ACK\u201d mechanism can be used for individual robot distinguishing other robots from obstacles or the boundary of arena"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003859_gt2013-95074-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003859_gt2013-95074-Figure4-1.png",
+        "caption": "Figure 4: Bochum test rig for large turbine bearings, technical drawing and front view",
+        "texts": [
+            " Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 08/12/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use 4 Copyright \u00a9 2013 by ASME 4.1 Mechanical design The Bochum test rig has been designed to examine large turbine bearings under practical operating conditions [1]. The nominal bearing diameter is 500 mm with a maximum length of 500 mm. Depending on the friction the driving power of 1.2 MW is sufficient to run the shaft at a top speed of 4000 rpm. The lubricant used is turbine oil ISO VG 32. Figure 4 shows the technical drawing of the test rig. The test bearing (1) is mounted in a rigid frame (9), which is connected with a pneumatic bellow (6). When pressurized, the bellow pulls the test bearing against the shaft (2) and thus creates a maximum bearing load of 1 MN. The force acting on the shaft is transferred to two symmetrically aligned support bearings (3). As well as the pneumatic bellow, these are supported by the test rig body (4). The hollow shaft is equipped with two piezoelectric pressure probes and two capacitive distance sensors, arranged with a circumferential distance of 90\u00b0 to each other",
+            " A slip ring transducer delivers the data to a PC. In addition to the stationary force, the test bearing can be loaded with sinusoidal forces. For this purpose, vibration generators (10) are attached to the frame of the test bearing. During the test procedure the relative movements between shaft and bearing are measured in order to determine the dynamic behavior of the test bearing. 4.1.1 Dynamic forces In addition to the stationary force, the test bearing can be loaded with sinusoidal forces. For this purpose vibration generators (10 in Figure 4) are attached to the frame of the test bearing in \u00b145\u00b0 to the vertical. Figure 5 shows the mechanical design of the vibrator in a section view. It consists of two pairs of imbalanced shafts (2 and 3), which are supported in a massive housing (1). These are connected with a toothed belt drive, so that looking in direction of the flange (12), the two adjacent shafts rotate in opposite directions. The shafts of each pair are precisely in phase to each other in such a way that the resulting force component of the vibrator, transverse to the radial bearing direction is zero"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure4-1.png",
+        "caption": "Fig. 4. The test rig layout and the arrangement of the detecting probes.",
+        "texts": [
+            " In order to avoid the premature problem of the GA optimization procedure, the computer simulation and fine tuning technique are introduced to the holobalancing method to ensure the optimal result. The amount and angular location of the correction weights can all be fine tuned and the balancing effect is simulated and displayed in real-time in the form of the 3Dholospectrum. Better scheme then could be found (if it is possible) without repeated test runs and the balancing efficiency is increased [5]. Furthermore, the balancing effect simulation of the correction scheme prior to its application reduces the balancing risk as well. A two mass rotor system test rig is shown in Fig. 4. This two mass rotor system is supported by oil-impregnated bronze bearings and connected to the motor with a flexible coupling. The rotor shaft has a nominal diameter of 10 mm, an overall length of 550 mm and a span between bearings of 325 mm. Two disks with a diameter of 75 mm are separated within the bearing span. Each disk has 16 screw holes, equally distributed on a circle with a radius of 32 mm, for adding trial weights. The rotating direction is counterclockwise when looking at the rig from the driving end"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003647_isie.2012.6237266-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003647_isie.2012.6237266-Figure1-1.png",
+        "caption": "Figure 1. Planetary gearbox with three planet gears: (a) Sun gear, (b) Ring gear, (c) Planet gears, (d) Planet-carrier.",
+        "texts": [
+            " Accordingly, in order to increase efficiency and also to decrease size and cost, planetary gearboxes are becoming more and more popular. Planetary gearboxes are comprised of arrangement of four different elements that produce a wide range of ratios in a compact layout. These four main components are: (a) Sun gear, an external gear located in the center of gearbox; (b) Ring gear, an internal gear, which has the same axe as sun gear; (c) Planet gears, the external gears mesh with sun and ring gears; (d) Planet-carrier, a support structure for planet gears (Fig. 1). By fixing one of the central elements (ring, planet-carrier or sun gear) and considering two other elements as the input and the output, several configurations would be achieved. For instance, in wind turbines normally the ring gear is the fix part, the planet-carrier is the input and the sun gear is the output, while in automatic transmission, there is a stage in which the ring gear is considering as the output gear. Despite the fact that planetary gearboxes are used in various industrial applications, their monitoring via electrical signature has not been considered in the literature"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003208_icra.2011.5979843-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003208_icra.2011.5979843-Figure1-1.png",
+        "caption": "Fig. 1. Sketch of the Model",
+        "texts": [
+            " Section II shows hybrid dynamic model of the walking gait built by Lagrange and conservation of angular momentum; Section III presents that the gait change is judged by eigenvalues of Jacobi matrix, which is solved by Newton-Raphson iterative method and based on Poincar\u00e9 mapping principle; Section IV analyzes the gait modification as the parameter c modifies, then we attain that, the gait enters chaos by period doubling bifurcation and obeys the law of all the period doubling bifurcation; Section V discusses the dynamic features of the gait when it enters the chaos by period doubling bifurcation as the slope increases, and analyzes the periodic rules of the chaotic gait; Section VI presents the conclusion of the above analysis, and proposes the suppression strategies or control strategies in terms of the bifurcation and the chaotic gait caused by different parameters of the robot. The robot in this paper is a 2-D walker with 4 straight legs. 2 medial legs and 2 lateral legs are connected respectively, which solves the problem of lateral instability. The robot I 978-1-61284-385-8/11/$26.00 \u00a92011 IEEE 2015 model is then in fact a biped walking robot model, and it is composed of two rigid straight legs and a passive hinge to connect legs, as Fig. 1 shows. All movements are limited in the plane of Fig. 1. Two straight legs of the model have the same mass and geometric parameters. Both legs are homogeneous. The mass of each leg is 1m , the moment of inertia relative to the centre of mass of each leg is J , the length is l , the distance between the centre of mass and the mass point of the hip is c . The mass of the hip is 2m ; to make the motivation more stable, the robot has arc feet, with the radius r . The model can be considered as straight-leg compass gait model. To reduce numbers of the parameters, also make the dynamics equation more universal, parameters are needed to be non-dimensional by setting r r l= , c c l= , 2 2 1m m m= , 2 1 1 1J J m l= , 3 3l l l= 2 2 2 1J J m l= and rescaled the time t (g/l) "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000809_robot.2008.4543556-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000809_robot.2008.4543556-Figure7-1.png",
+        "caption": "Fig. 7. Rover test bed",
+        "texts": [
+            " The performance of the proposed control is evaluated based on the distance and orientation errors. Fig. 6 shows the overview of the experimental setup with our rover test bed on the tiltable test field. The test field consists of a flat rectangular soil-vessel in the size of 2.0 by 1.0 [m]. The vessel is filled up with 8.0 [cm] depth of a Toyoura Sand, which is loose sand and standard sand for terramechanics research filed. The vessel can be inclined up to 20 [deg]. 1) Rover test bed: The four-wheeled rover test bed as shown in Fig. 7 has a dimension of 0.44 [m] (wheelbase) \u00d7 0.30 [m] (tread) \u00d7 0.30 [m] (height) and weights about 13.5 [kg] in total. Each wheel of the rover has a diameter of 10 [cm] and a width of 6.4 [cm], and it is covered with paddles having a height of 0.5 [cm]. Every wheel can steer in the range of \u00b1 90 [deg]. An on-board computer located inside of the rover executes all steering and driving maneuvers based on the proposed slip compensation control. 2) Measurement system for velocity and orientation: Measurement of the slip motions is the key procedure to achieve the proposed control"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000207_s00542-006-0309-6-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000207_s00542-006-0309-6-Figure11-1.png",
+        "caption": "Fig. 11 Finite element model of 1-in. disk drive using thick enclosure",
+        "texts": [
+            "3 Head-slap shock simulation In the next step of the study, the two HDD models with spindle and disk were used to study the effect of external shock. A half-sine acceleration of 700 G amplitude [1 G = acceleration of gravity (9.81 m/s2)] and 1 ms duration was applied to the base (enclosure) of the 1-in. disk drive model. The time step used for the head slap simulation is 20 ls. A simulation time of 10 ms was used which gives a total of 501 time steps. This simulation was for non-operational linear shock condition. The transient analysis of the drive response to a shock was performed using a non-linear finite element solver (LS-Dyna 2006). Figure 11 shows the model using the thick enclosure with disk, spindle, arm, suspension and the head gimbal assembly. Contact was defined between components to avoid interference. The contact algorithm used is a surface-to-surface contact with one surface being assigned as \u2018\u2018master\u2019\u2019 surface and the other surface being a \u2018\u2018slave\u2019\u2019 surface. This algorithm ensures that the slave elements do not penetrate the master elements by using a penalty method. Figure 12 shows the relative displacement between the slider and the disk due to a shock load in the z-direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001580_1.2844956-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001580_1.2844956-Figure1-1.png",
+        "caption": "FIGURE 1. The Device of the Container. I - Electric Coils of the Engine of a Gyroscope, 2 - a Massive Cylindrical Part of a Rotor, 3 - the Case of the First Gyroscope, 4 - Plugs of Power Supplies of Engines of Gyroscopes, 5 - the Case of the Second Gyroscope (it is Shown without a Section), 6 - the Case of the Container.",
+        "texts": [],
+        "surrounding_texts": [
+            "The simple way of estimation of difference value (ap - a^) is based on weighing of a rotor of a mechanical gyroscope with a horizontal axis of rotation. The role of elastic forces is played here by centripetal forces. It is possible to show that the weight P of the rotor in form of a cyhnder with radiuses Ri and R2 is equal to P=Mg, l - ( \u00ab p - \u00ab c ) 3ng,{Rl-R^) CO (4) where co is the angular speed of rotation. Such an experiment was executed in 1999-2000 at Saint Petersburg (Dmitriev and Snegov, 2001). In this case, we weighed a pair of coaxial rotors rotating in opposite directions for compensation of the total angular moment of the container (Ri=I5 mm, R2=25 mm, M=250 g), as shown in Fig. I. The obtained experimental dependence is shown in Fig. 2. At a speed of rotation of 18.6 thousand rev/min the relative reduction of weight of a rotor was equal to 3-10\"''. The estimated value of (ap - a j is near to 10\"'. The factor a\u0302 , alone was evaluated by precision measurements for the restitution coefficients of an elastic impact of a ball against a massive metal plate. In these experiments a plate (and a ball trajectory) took horizontal and vertical positions (Dmitriev, 2002). Acceleration of the ball during impact duration exceeded of W-go. The difference of restitution coefficients in vertical (ki) and horizontal (k2) quasi-elastic impacts of the ball is shown in Fig. 3. The Order of Value of the Factor a^ can be Estimated by the Formula a. 2 1 1 + A:, (5) The speed of the ball before impact is about 3.5 m/s, which gives a lad^lO'^ that is unexpectedly big. Interesting results obtained by M. Tajmar's group in experiments with a rotating superconductor (Tajmar et al., 2007) probably have a physical nature close to the one discussed in the present work. MEASUREMENTS OF TEMPERATURE DEPENDENCE OF BODY WEIGHT If indeed there is an influence of acceleration of elastic (electromagnetic in nature) forces on gravitation, then there will be the temperature dependence of body weights due to the thermal movements inside the body. The acceleration of microparticles in their thermal movement directly depend on their energy, and therefore from the absolute temperature of body. It is possible to show that in a classical approximation at temperatures higher than the Debyetemperature, the temperature dependence of body weight is described by the formula P = Mg, ^ 0 (6) where C is a the factor dependent on physical characteristics (including density and elasticity) of bodies and T is the absolute temperature. According to Equ. (6), an increase in the absolute body's temperature will cause a reduction of its weight. Such an effect was indeed observed in exact weighing of metal samples from nonmagnetic materials heated with ultrasound (Dmitriev, Nikushchenko and Snegov, 2003). The layout of the hermetically sealed container shown in Fig. 4. An example of the experimental dependence of a sample weight in the process of its heating and cooling is shown in Fig. 5. 2 0 ^ -2 I -6 I -8 5-10 (/3 ^-12 -14 -16 0 2 A ^^ 6 1 8 10 12 A A A A A A ' ' ^ A A \\ i ^ ^ A ^ ^ ' A A A ^ A A -"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001675_j19660001059-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001675_j19660001059-Figure1-1.png",
+        "caption": "FIGURE 1",
+        "texts": [
+            " Pu bl is he d on 0 1 Ja nu ar y 19 66 . D ow nl oa de d by C or ne ll U ni ve rs ity L ib ra ry o n 28 /0 7/ 20 16 1 3: 04 :2 9. Inorg. Phys. Theor. q = p + 1 t = t + l solve equations D and a bv matrix inver- --q+q+1 1061 Scheme 2 Stage C t = O p = o sion; find Kl(t), and wkighting factors WlW. W*(t, No I q = n ? No I p = (n - l )? lye, Stage E for each value of 1/K, and Kz were derived from the premise that the best values of a constant are obtained when the two lines intersect a t 90\u201d (see Figure 1). To eliminate the disparity in the scales of the two constants and the corresponding intercepts on the axes, we put x = y = 1 and modified the intercepts accordingly. The angle between the lines is then given by 8 = arc tan(Y(,jy - 1) - arc tan(Y@>/y - 1) (8) A function is required which will be unity when 8 = 90\u201d; this is evidently the modulus of sine 8. To improve the discrimination we used Wy = exp[((sin 81 - l)/lsin 811 (9) Equation (9) gives Wy = 1 when 8 = 90\u201d. W , was set equal to zero for intersections giving In stage E, the first average values of the constants negative constants"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003559_s12666-012-0151-8-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003559_s12666-012-0151-8-Figure4-1.png",
+        "caption": "Fig. 4 Computational domain and grid structure for cylindrical keyhole with flat bottom",
+        "texts": [
+            " Shrinking of the heat source from the top surface towards the bottom in case of conical keyhole might be the cause. Subsequently, a cylindrical keyhole is thought to be a better substitute for correcting the mismatch both at the top and bottom of the fusion line. The literature also suggests that the keyhole shape can be taken to be cylindrical if the keyhole is deep [2, 13]. It is expected that bulging of the fusion line near the bottom can be better simulated with a cylindrical keyhole than a conical one where the strength of the heat source diminishes near the bottom. Figure 4 shows the computational domain and grid structure used for the simulation with cylindrical keyhole with flat bottom. Figure 5 shows comparison between the simulated and experimental profile. It can be seen here that although the fusion line matching has improved at the bottom, it has deteriorated at the middle. Besides, the predicted thickness of fusion zone at the bottom has superseded the actual thickness. It is a simpler variation of the previous model. In this model all the parameters are same as that of Case B except the shape of keyhole at the bottom"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003061_004051756503500913-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003061_004051756503500913-Figure1-1.png",
+        "caption": "Fig. 1. Recorder traces of the 2nd and 8th compressive and expansion cycles.",
+        "texts": [],
+        "surrounding_texts": [
+            "8. Simpson. J., Callegan, A. T., and Sens. C. L., Textile Ind. (11): 101, 104-105, 108-109, 115 (1960). 9. Temmerman, R. and Hermanne, L., \"Application of the Index of Irregularity to the Study of Spinning on the Cotton Systems,\" J. Textile Inst. 41, No. 11, T411-421 (1950).\nJack Simpson Southern Utilization Research and Development Division 1100 Robert E. Lee Boulevard New Orleans. Louisiana\nFactors Influencing the Resistance to Compression of Wool Fiber Ensembles\nMay 19, 1965 \u2019\nDear Sir:\nA piston-and-plunger method, similar to those already described in the literature [ Z, 3, 5, 6. 7 ],, has been used for the determination of the bulk compressibility of a wool sample. The method has been proved consistent and two applications of it are given. The first application, that of fiber attributes, has been thoroughly studied by Van Wyk [10] who has also reviewed earlier literature. It was necessary B0 repeat this part of his work\n\u2019\nso as to eliminate the influence of these attributes from\nany subsequent work to be done with this method. The second, that of natural and artificial weathering, emphasizes the differences between these two forms of\nweathering.\n\u2019 Experimental Procedure \u2019\n. Cape Merino wool was used, with the tip half-inch of the staple removed. The root portions were cleaned by the standard method of washing at room temperature in a succession of bowls of petroleum ether. 0.1 % Lissapol , NX solution, alcohol, and ether. The wool was sub-\nsequently conditioned at 65% RH and 20\u00b0 C, after which it was teased out with hand cards. All vegetable matter was removed, and samples of 2.5 g each were weighed out to the nearest 5 mg. This sample size coincides with that used in the airflow meter for the diameter determination, which was always done after the compression tests. A weighed sample was placed in a container 4.6 cm in height and\u2019 5.25 cm (2 in.) in diameter. In lieu of a compression cell, an arrangement was made with a cylindrical container suspended from the top jaw of an , Instron Tensile Tester (Table Model), so that the bot. tom jaw or movable cross head was perpendicularly\nabove it. A plunger of 5.05-cm diameter was fixed to this cross head and was used for compressing the wool. With this arrangement all the speeds and load ranges of the Instron could be applied. The wool underwent compression-expansion cycles between two points 1 cm (equivalent to a wool density of 88 mg/cc) and 3 cm (29.3 mg/cc) from the bottom of the cylinder. A reading of the force (expressed in g/cm2) was taken on the eighth compressive cyde reg-\nistered on the Instron at a compression point 1.5 cm from the bottom of the cylinder (woot density Scl.7 nig/cc -), which was then designated the rcsislanct\u2019 to\ncontfiression. Initially, eight cycles were chosen hecause the saml>le had to come to some form of equilibrium with successive hysteresis cycles (differing by less than Z~~,. It was found, though, that an excellent correlation (0.99) existed between force readings taken at the eighth and second cycles (points li and F. I-ig. 1 ) ; according\u2019ly, only twQ cycles per test were thereafter done.\nRepeated testing of a sample caused its resistance to compression to increase due to a slow felting o,f the woot [10]. To avoid this, no sample was used more than\nat University of British Columbia Library on June 23, 2015trj.sagepub.comDownloaded from",
+            "857\nthree times, with teasing out of the wool hy hand between successive tests.\nThis experimental method appeared to he consistent, as the cuelfiicient of variation cm results obtained on different samples hy different operators was of the order of 4~~;. Contrary to the results obtained by llees [9]. friction against the sides of the compression cylinder was found not to contribute to the resistance to compression : this was shown in a series of experiments wherein the side-wans and the wool sample were either washed in ether or alternatively lubricated with class A combine oil. Bu consistent trend could he detected.\nI\u2019Irc\u00b7 IuJTrmmu\u00b7 of I-iber I_unyllr, l)imm\u00b7tur, ami C~rinr/c\nFiber length showed a small. but nevertheless sigiiificant, influence. \u2019I\u2019he decrement in the resistance to\ncompression of O.UJ g cm\u2019= per cm of fil>E\u00b7r length. however. may 1e considered negligihly small in comparison to the irtfluence of other attributes. The effect of fiber diameter and of staple crimp were assessed cm a total of 81 fleeces divided into five lots : for two of these lots (A and It ), the fleeces came from various parts of the country, whereas lots C. 1 ), and F were each composed of fleeces from single Hocks only. so that the wool within a lut had heen derived from\nsheep subjected to a more uniform treatment and feeding. Partial regression analysis indicated that the reKression coefficients of resistance to compression against sample diameter and crimp were generally significant at ill&dquo; or higher levels. The standard errors of the lots reflected. to a certain extent, cm the origin of the woo! as the figtires were smaller for tots C. I >, and K than for lots A and li.\nTtte v:~ri:~l~le diameter X crimp \u2019cm (1()J was also applied and proved useful for assessing the data with a single variable. For instance, a covariance analysis ltetween two sets of twclve samples, drawn randomly from lots :B and C. with the above variable, g-ave a siKnificant difference (F==8.6. 1 and 21 decrees of freedom) between the lots. This illustrates once more the importance of the origin of the wool. even when crimp and diameter are taken into consideration. 1 he distribution of the points is illustrated in Figure 2 where data for samples from lots C, I), and E are set out against this variable. This compound variable also seems to give a better regression, as the correlation coefficient for it was\ngenerally near 0.75 in contrast to the total correlation coetiicients of the linear regression which was of the order of 0.6.\nh, flll ell cc of lt\u2019ctithering \u2019\nTwo forms of weathering were studied, viz., natural weathering and exposure of selected wool samples to a xenon arc lamp. The method employed in the former was to use all the tips discarded in the foregoing exheriments and, after preparing them in the manner set out before, their resistance to compression was determined. Some of the cut tips consisted of perhaps 50% relatively\nf\n~\nvariable crimp/cm X diameter (micron).\nunweathered wool. while the crimp of some samples was undeterminable. so that from the outset this part of the experiment was viewed with some doubt. No consistent trend could he observed for the influence of natural\nweathering on the resistance to compression of wool. On the other hand, artificial weathering showed a most pronounced effect. After an initial test to determine the resistance to compression, the wool, in its opened-out form, was exposed to a 6000-watt xenon arc lamp (Atlas Weather-0-Meter) in a 30~ RH atmosphere. Three samples were exposed for each of eight different periods ranging from 1 hr to 24 hr. The resistance to compression increases with exposure (Fig. 3), but, after seven or eight hours, it levels off at a value about 19% higher than for the original wool. The results in this section necessitated a careful control ex-\nperiment which consisted of allowing certain samples to undergo the complete testing procedure as set out before; except that the xenon light of the Weather-O-Meter was not switched on. This was necessary, as it was thought that the vibration of the rotating frame and the circulation of air could have a felting effect, with a consequent rise in the resistance to compression. An increase of 4% over the value for the control was recorded, leaving\nat University of British Columbia Library on June 23, 2015trj.sagepub.comDownloaded from"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003080_jmer.9000033-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003080_jmer.9000033-Figure2-1.png",
+        "caption": "Figure 2. Slider crank mechanism.",
+        "texts": [
+            " At first stage the combustion chamber pressure curve of Nissan Z24 engine was measured in MegaMotor's power test lab (Engine, Gearbox and Axel Manufacturing located in Tehran province in Iran, info@megamotor.ir). These experimental data have been shown in Figure 1. Because of different type of motion in this mechanism; such as: linear, linear-rotation and rotation, there is inertial force in the system. The inertial force has important role in engine slider-crank mechanism, so behavior of this force must be known. For analyzing of inertia force, the kinematics of mechanism should be defined. Kinematics analysis of slider-crank mechanism The engine slider-crank mechanism has been shown in Figure 2. The piston has linear motion in x direction in this figure: )cos()cos( \u03b2\u03b8 lrx += (1) Where, r is the crank radius, L is the connecting rod length, \u03b8 is the crank rotation angle and \u03b2 is the connecting rod angle with x axis. From Figure 1, one can obtain that: )sin(.)sin(. \u03b2\u03b8 lr = (2) l r n = (3) And thus: )(sin1)cos( 22 \u03b8\u03b8 nlrx \u2212+= (4) Using Taylor series in Equation 4: ...)sin( 16 )sin( 8 )sin( 2 1 1 ...))sin(( 16 1 ))sin(( 8 1 ))sin(.( 2 1 1))(sin(1 6 6 4 4 22 64222 +\u2212+\u2212= +\u2212+\u2212=\u2212 \u03b8\u03b8\u03b8 \u03b8\u03b8\u03b8\u03b8 nn n nnnn (5) Because n is less than 1(about 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure4-1.png",
+        "caption": "Fig. 4 Kinematic chain of the McGill SMG: front view of the home configuration, with all parallelograms in the YZ plane",
+        "texts": [
+            " In the case of Scho\u0308nflies motions, two components of the angular velocity vanish, and hence a fourdimensional twist array suffices, namely* t~ v _pT T \u00f01\u00de The kinematic relations between actuated joint rates and Cartesian velocities are now derived. The geometric relations between the actuated joint variables and the pose of the MP are not derived here, their derivation, as pertaining to an evolved version of the MP, being available in reference [15]. By virtue of Lemma 1, the velocities of the points OI3 and OII3, shown in Fig. 4, are equal to the velocities of the points OI4 and OII4 respectively. The two position vectors pI and pII of OI3 and OII3 respectively are next introduced pI~aI1zaI2zaI3, pII~l0jzaII1zaII2zaII3 \u00f02\u00de Differentiation of both sides of equation (2) leads to k|rJ\u00f0 \u00de _hJ1z f J|aJ2 _hJ2z f J|aJ3 _hJ3~ _pJ , J~I , II \u00f03\u00de where the relations below have been used rJ~aJ1zaJ2zaJ3, _rJ~ _aJ1z _aJ2z _aJ3 \u00f04a\u00de _aJ1~ _hJ1k|aJ1, _aJ2~ _hJ1kz _hJ2f J |aJ2, _aJ3~ _hJ1kz _hJ3f J |aJ3 \u00f04b\u00de * As the proof is straightforward, it is not included here",
+            " Equation (3) includes actuated and passive joint rates, the latter being eliminated by cross-multiplying both sides of equation (3) by (fJ6aJ3), for J5 I, II, thereby obtaining f J|aJ3 | k|rJ\u00f0 \u00de _hJ1z f J|aJ3 | f J|aJ2 _hJ2 ~ f J|aJ3 | _pJ \u00f05\u00de Let vJ: f J|aJ3 | k|rJ\u00f0 \u00de : f J|aJ3 T rJ h i k{ f J|aJ3 T k h i rJ \u00f06\u00de and DJ: f J|aJ3 T aJ2 \u00f07\u00de Furthermore, by virtue of the Scho\u0308nflies motion undergone by the MP _pJ~ _pz 1 2 _wsJk| pI{pII\u00f0 \u00de \u00f08\u00de with p\u0307 denoting the velocity of both the operation point P and of the midpoint P9, of position vector p9, of OI3OII3, these points being shown in Fig. 4. Moreover, sJ~ z1 if J~I {1 if J~II \u00f09\u00de Substitution of equations (6), (7), and (8) into equation (5) yields vJ _hJ1zDJ f J _hJ2 ~ f J|aJ3 | _pz 1 2 _wsJk| pI{pII\u00f0 \u00de \u00f010\u00de Now, in order to rewrite equation (10) in a compact form, additional definitions are introduced AJ: 1 2 sJWJ k| pI{pII\u00f0 \u00de\u00bd WJ [ R3|4 \u00f011a\u00de BJ: vJ DJ f J [ R3|2 \u00f011b\u00de _hJ: _hJ1 _hJ2 \" # , t: _w _p \" # , wJ:f J|aJ3 \u00f011c\u00de where WJ is the cross-product matrix* [14] of vector wJ. With the above definitions, equation (10) becomes AJ t~BJ _hJ , J~I , II \u00f012\u00de Now, upon assembling the two kinematic relations displayed in equation (12), the kinematic model of the whole system is obtained, namely At~B _h \u00f013\u00de in which h\u0307 is the four-dimensional vector of actuated joint rates and A and B are the 664 forward and inverse kinematics Jacobians, defined as A: AI AII \" # [ R6|4, B: BI O32 O32 BII \" # [ R6|4, _h: _hI _hII \" # [ R4 \u00f014\u00de with O32 denoting the 362 zero matrix",
+            " What this means is that the gear axis A9 is one of the principal axes of inertia about C9, the other two principal axes being any pair of mutually orthogonal axes passing through C9 as well as lying in a plane normal to A9. The three principal moments of inertia of the bevel gears are thus the axial moment of inertia I9a, the other two being identical and referred to as the transversal moment of inertia I9t. Furthermore, the angular velocity of the gears can be decomposed about two mutually orthogonal components, one parallel to A9, the spin, the other vertical, the precession. These items are given below v9~vPkzvt fJ , _c9~v9|p9, p9:c9{aJ \u00f049\u00de where fJ was introduced in Fig. 4, c9 is the position vector of C9, which lies along A9, and aJ is the position vector of a point of AJ , for J5 I, II. Moreover, it is assumed, without loss of generality, that the foregoing point is chosen so that p9 is horizontal, i.e. pT 9k~0. Furthermore, the norm ||k6p9|| will be needed presently, whose square is k|p9 2~ kk k2 p9 2sin2 k, p9 ~ p9 2~r29 \u00f050\u00de Upon substitution of vp and vt in terms of motor rates, t9 can be expressed, for each limb, as t9J~ 2 D {kz N r6,5 f J {Nk{ N r6,5 f J k|p9 k|p9 \" # |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl} S9J _qJ , J~I, II \u00f051\u00de where factor 2 is included in order to account for the two horizontal bevel gears of each limb"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure11-1.png",
+        "caption": "Fig. 11. coupled system for global torsion (modal approach)",
+        "texts": [],
+        "surrounding_texts": [
+            "The previous described FBS approach makes it theoretically possible to extract static stiffness values from free-free FRF measurements. This is however only one value: it doesn't give any information on the modal contribution of the flexible car body modes to the global static stiffness. The FBS approach could easily be extended to make it possible to retrieve the modal contribution of the flexible car body modes. A full free-free modal identification could be performed on the measured FRF's. From this modal base, the original FRF's can be synthesized using equation (4) (4) The same FBS approach can now be used but instead of working with the measured FRF's describing the vehicle body subsystem, the synthesized FRF's will be used. Of course a modal truncation error will be made by only taking a limited modal base into account. This effect is very important at the clamping and loading locations. To compensate for the missing flexibility, the residual stiffness at the clamping and loading points can be integrated in the FBS equation. Instead of connecting subsystem A (vehicle body) rigidly to subsystem B (ground), the two subsystems are now connected by a spring representing the residual stiffness as shown in Fig. 10 and 11. This residual stiffness can be estimated based on the upper residual terms obtained from the modal identification of the system under free-free conditions. If one is interested in only the contribution of a subset of modes to the global stiffness, one can limit the FRF synthesis (equation (4)) only to that particular subset of modes. In Fig. 12 the torsional stiffness is plotted over an increasing number of modes taken into account during the FRF synthesis. Large drops indicate modes with a large contribution to the global torsional stiffness. The first torsion mode is of course an important contributor. But also the residual stiffness at the four dome points has a significant contribution towards the global torsional stiffness. <figure 12 here>"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002470_s1758825111001093-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002470_s1758825111001093-Figure1-1.png",
+        "caption": "Fig. 1. Investigate model of the angular misalignment.",
+        "texts": [
+            " Furthermore, we determined stiffness and damping matrices in the first step, while we present the adopted theoretical model as well as the different results in the second. Rotor misalignment determining a machine\u2019s reliability is a condition wherein the driving and driven machine shafts are not on the same centerline. To accommodate In t. J. A pp l. M ec ha ni cs 2 01 1. 03 :4 91 -5 05 . D ow nl oa de d fr om w w w .w or ld sc ie nt if ic .c om by U N IV E R SI T Y O F L IV E R PO O L o n 11 /2 5/ 15 . F or p er so na l u se o nl y. this fault that can severely damage bearings and gear, flexible couplings are used. Figure 1 shows the simplified theoretical model investigated by Slim et al. [2007] and that illustrates the angular misalignment. The model hypotheses are the following: rigid shafts with constant inertia, the coupling consisting of a ball-and-socket joint out of line of the axes intersection, the motor shaft (1) is carried by a rigid bearing (3) have only one rotation around its axis. A rigid weighing rotor (2) carried by two AMBs (4) and (5), and misaligned of an angle \u03b1(t) + \u03b1E in (x, y)-plane. \u03b1E is an imposed angle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003045_s11044-012-9326-7-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003045_s11044-012-9326-7-Figure4-1.png",
+        "caption": "Fig. 4 Asymmetric impact posture",
+        "texts": [
+            " eight-legged, and l0 = 1.0 [m]. Figure 3 shows the essential part of the TRW which determines the stance-phase motion. This is just a variable-length pendulum and its dynamic equation becomes very simple. Let q = [\u03b8 l ]T be the generalized coordinate vector; the dynamic equation of the TRW then becomes where M(q) = [ Ml2 0 0 M ] , h(q, q\u0307) = [ 2Mll\u0307\u03b8\u0307 \u2212 Mlg sin \u03b8 \u2212Ml\u03b8\u03072 + Mg cos \u03b8 ] , S = [ 0 1 ] u, and u [N] is the control force of the linear actuator to extend/retract the telescopic-leg frame. Figure 4 shows the configuration of impact phase. From this figure, the relations of the angular position and the stance-leg length can be found to be \u03b8+ = \u03b8\u2212 \u2212 \u03b1, (2) l+ = l0. (3) The point H in the figure is the ground projection of the CoM, Xg , and Xg is the step length which is identical to the distance Xg moved during one step. The following derives the transition equations of the velocities. We introduce an extended coordinate vector q\u0304 = [ x z \u03b8 ]T where (x, z) is the stance-leg\u2019s tip position. The inelastic collision is then modeled as M\u0304(\u03b8) \u02d9\u0304q+ = M\u0304(\u03b8) \u02d9\u0304q\u2212 \u2212 J I (\u03b8)T\u03bbI , (4) J I (\u03b8) \u02d9\u0304q+ = 02\u00d71, (5) where M\u0304(\u03b8) \u2208 R 3\u00d73 is detailed as M\u0304(\u03b8) = \u23a1 \u23a3 M 0 Ml1 cos \u03b8 0 M \u2212Ml1 sin \u03b8 Ml1 cos \u03b8 \u2212Ml1 sin \u03b8 Ml2 1 \u23a4 \u23a6 . (6) Note that \u03b8 in this case is not that in q but that in q\u0304 , and is equal to \u03b8\u2212 in Fig. 4. The extended coordinate does not take stance-leg exchange into account, that is, q\u0304\u2212 = q\u0304+ and \u03b8\u2212 = \u03b8+ hold in this system. We must reset \u03b8+ for q following Eq. (2) after impact. J I (\u03b8) \u2208 R 2\u00d73 is the Jacobian matrix for inelastic collision. We assume that there is a high friction at the contact point and the leg\u2019s end-point does not slide immediately after impact. The conditions for velocity constraint are given by d dt ( x + l1 sin \u03b8 + l0 sin(\u03b1 \u2212 \u03b8) )+ = 0, (7) d dt ( z + l1 cos \u03b8 \u2212 l0 cos(\u03b1 \u2212 \u03b8) )+ = 0",
+            " (15) By extracting the third element of \u02d9\u0304q+, we get \u03b8\u0307+ = I + Ml0l1 cos\u03b1 I + Ml2 0 \u03b8\u0307\u2212. (16) We then obtain lim I\u21920 \u03b8\u0307+ = l1 cos\u03b1 l0 \u03b8\u0307\u2212. (17) Further, by setting l0 = l1, we obtain \u03b8\u0307+ = \u03b8\u0307\u2212 cos\u03b1. (18) This equation is the same as that of the rimless wheel [9] and the simplest walking model [10]. On the other hand, from the geometric relation at impact, the relation l1 cos \u03b8 = l0 cos(\u03b1 \u2212 \u03b8) holds and the first element of \u02d9\u0304q+ is found to be zero, that is, x\u0307+ = 0. In addition, l1 sin \u03b8 \u2212 l0 sin(\u03b8 \u2212 \u03b1) is identical to the step length, Xg , in Fig. 4, and the second element of \u02d9\u0304q+ is thus always positive. As is seen from Fig. 4, the velocity immediately before impact is V \u2212 = l1\u03b8\u0307 \u2212 [m/s] and that immediately after impact is V + = l0\u03b8\u0307 + = l1\u03b8\u0307 \u2212 cos\u03b1 [m]. The kinetic energies immediately before and immediately after impact then become: K\u2212 = 1 2 M ( V \u2212)2 = 1 2 M ( l1\u03b8\u0307 \u2212)2 , (19) K+ = 1 2 M ( V +)2 = 1 2 M ( l1\u03b8\u0307 \u2212 cos\u03b1 )2 . (20) The energy-loss coefficient then becomes \u03b5 = K+ K\u2212 = cos2 \u03b1 (21) regardless of the ratio of leg length. In a passive rimless wheel, the impact posture is always identical, and both the energyloss coefficient and the restored mechanical energy are kept constant automatically",
+            "4 Efficiency analysis Let us introduce criterion functions before performing numerical analysis. Let T [s] be the steady step period. For simplicity, in the following, every time immediately after impact is being denoted as 0+ and every time immediately before impact as T \u2212 by resetting the absolute time at every transition instant. Thus T + means the same as 0+. The average walking speed is then defined as v := 1 T \u222b T \u2212 0+ X\u0307g dt = Xg T , (28) where Xg [m] is the X-position at the CoM and Xg := Xg(T \u2212) \u2212 Xg(0+) [m] is equal to the step length shown in Fig. 4. The average input power is defined as p := 1 T \u222b T \u2212 0+ |l\u0307u|dt. (29) Energy efficiency is then evaluated by specific resistance (SR): SR := p Mgv , (30) which means the expenditure of energy per unit mass and per unit length, and this is a dimensionless quantity [3, 4]. The main question of how to attain the energy-efficient locomotion rests on how to increase the walking speed v while keeping p small. Let E [J] be the restored mechanical energy per step; then the following magnitude relation holds: p \u2265 1 T \u222b T \u2212 0+ l\u0307udt = E T ",
+            " This is because the CoM is strongly stalled around the potential barrier. From Figs. 9(b) and (d), we can see that the walking speed and the restored mechanical energy monotonically increase as s increases for all Tset. These imply that the forward motion is accelerated monotonically with respect to the extended leg length. From Fig. 9(c), however, we can see that the energy efficiency worsens in return for it. E monotonically increases with the increase of s without generating the negative input power, whereas the step length in Fig. 4 satisfies X2 g = l2 0 + l2 1 \u2212 2l0l1 cos\u03b1 and its derivative with respect to l1 becomes \u2202 X2 g \u2202l1 = 2l0(s \u2212 cos\u03b1) > 0 (33) because of s > 1. Therefore, Xg also monotonically increases with the increase of s. We then conclude that the increasing rate of E is much larger than that of Xg . This section extends our method to a planar telescopic-legged biped robot and discusses the role of asymmetric shape of human foot from the point of view of ankle-brake effect. 3.1 Modeling of telescopic-legged biped robot Figure 10 shows the planar telescopic-legged biped model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001559_acemp.2007.4510568-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001559_acemp.2007.4510568-Figure3-1.png",
+        "caption": "Fig. 3. 2D flux path",
+        "texts": [
+            " In this paper, a 3D magnetic equivalent circuit is proposed and its validity is established by comparing the results it gave to those given by a finite-element model. In addition to the polar symmetry, the homopolar structure has a plan of symmetry. This plan is noted xOy and illustrated in Fig.1. Consequently, twelfth of the structure is enough in order to study the machine\u2019s behaviour (Fig.2). A. 2D flux path 2D flux path consists of the path of the part of the flux created by permanent magnets that participate to the energy conversion (Fig. 3). Its modelling is shown in Fig. 4. Reluctances appearing in the model are indicated in the nomenclature. Fig. 4. Magnetic equivalent circuit for 2D\u2019s path Where : a\u03b5 is the magnetomotive force produced by permanent magnets : 0\u00b5 \u03b5 aerB a \u00d7 = ea is the permanent magnet thickness Br is the residual flux density The total excitation flux is given by : ( ) ( )( ) '22' ' 2 1212341212 1212 RRRRRRR RR ++++ + \u00d7= eqa a v p \u03b5 \u03d5 (1) The reluctances appearing in the previous expression are calculated from the dimensions and the magnetic permeability of each part of the machine [7] and [8] : ( )\u222b \u00d7 = 2 1 l l lS dl \u00b5 R (2) Where : \u00b5 is the magnetic permeability of the concerned part of the circuit l is the length of the flux\u2019s path S is the cross-sectional area of the path The calculation of the reluctance of a half rotor pole is given as an exemple"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000469_14644193jmbd78-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000469_14644193jmbd78-Figure8-1.png",
+        "caption": "Fig. 8 Forces acting on piston skirt",
+        "texts": [
+            " 9) h0 = xL \u2212 a sin \u03d1 \u2212 rtt cos \u03d1 (17) For simplicity, it is considered that only four corners of the piston touch the cylinder wall when the piston is in operation (due to its secondary motions, lateral and tilting) and these four positions are shown numbered in Figs 8 and 9 (points 1 to 4). This approach follows that of Haddad and Howard [35], who used the four corner model with spring-damper elements, which are clearly not representative of tribological conditions. Force component due to pressure variation is orthogonal to the piston surface at these four corners of the piston as shown in Fig. 8. It should be noted that at the extremities of the cycle (at TDC and BDC) due to cessation of lubricant entraining motion (reversal in sliding velocity), any lubricant film is retained by squeeze film action or trapped in between asperities of the contiguous surfaces. It is generally accepted that a mixed regime of lubrication occurs in such locations, where asperity interactions occur. The friction force in such instances is a combination of viscous action, described Proc. IMechE Vol. 221 Part K: J"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003906_ssd.2012.6198073-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003906_ssd.2012.6198073-Figure2-1.png",
+        "caption": "Figure 2. Schematic of the removable device mounted on a medical ultrasound probe with details. Frame spare parts (1a, 1b, 1c, 1d); strain gauges load cell (2)",
+        "texts": [
+            " The most important contact force component for probe control is the reaction force normal to the body surface. However, for soft tissues exploration, too high contact force of the ultrasound probe tends to deform the internal structure of the patient\u2019s tissue thus distorting the diagnosis. The contact force value varies from 5 to 20N [1]. Ultrasound images of a dog\u2019s ladder at three levels of force: (a) low, (b) suitable and (c) high are shown in Fig. 1. 978-1-4673-1591-3/12/$31.00 \u00a92012 IEEE II. GENERAL DESCRIPTION Fig. 2 shows a schematic of the removable device mounted on a medical ultrasound probe. The device consists of a shell which is a mechanical part with the same shape as the ultrasound probe and slightly larger than the probe. It can slide along the probe as the physician applies a longitudinal force. Bearings have been mounted at the shell base in order to reduce frictions with the probe. A bracket with two orthogonal arms was welded to the bottom of the shell and a strain gauges load cell was mechanically coupled to the probe on one side and to the bracket on the other side using clamps"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003977_j.proeng.2012.09.561-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003977_j.proeng.2012.09.561-Figure12-1.png",
+        "caption": "Fig. 12 The fourth modal shape in Solid Works",
+        "texts": [],
+        "surrounding_texts": [
+            "The aim of the measurement was to find out natural frequencies and modal shapes of pipeline model for two cases. The first case when springs of elastic support were not preloaded and when they were. Measurement results comparison acquired by Pulse6 for the case when springs were not preload with the results acquired by software Solid Works are presented on next figures (fig.5 - fig.14). Preload of pipeline support springs was created in one half of their length and the second half conserved its previous stiffness. Modal shapes did not differ for both investigated cases but differ in natural frequencies. From this reason we present only natural shapes of the first case, without preload. As we can see from the images there is from fourth modal shape also a deformation of pipeline cross-section and not only distortion of its shape. Corresponding natural frequencies of the particular modal shapes are presented in a table 1. Table 1. Natural frequencies of corresponding modal shapes Natural frequency no. 1. 2. 3. 4. 5. MTC Hammer 20,5Hz 36Hz 82Hz 263Hz 350Hz MTC Hammer P 25Hz 42Hz 85Hz 263Hz 352Hz Solid Works 29,8Hz 35,5 86,3Hz 282Hz 374Hz Where MTC Hammer means the measurement of pipeline vibration by system Pulse when the springs were not preloaded and MTC Hammer P means the measurement of pipeline vibration by system Pulse when the springs were preloaded. The results acquired by software SolidWorks are for the case when springs were not preloaded."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003779_s11071-012-0647-0-Figure14-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003779_s11071-012-0647-0-Figure14-1.png",
+        "caption": "Fig. 14 Parallel double inverted pendulum system; two inverted pendulums are pinned to a moving cart",
+        "texts": [
+            "1 System dynamics It is impossible to control the parallel double inverted pendulum systems with the same natural frequency of the pendulums. By changing the lengths of the pendulums, we can overcome this problem. By taking the cart movements into consideration, and its intensive effects especially on the shorter pendulum with higher frequency, it is accepted that the parallel inverted pendulum systems are the most difficult to be stabilized in the inverted pendulums category [51]. As shown in Fig. 14, the parallel double inverted pendulum system consists of a cart moving on a rail, a longer pendulum 1 hinged on the right side of the cart, and a shorter pendulum 2 hinged on the left side of the cart. An input force is driving the cart. The parameters M , m1, and m2 are the masses of the cart, pendulum 1, and pendulum 2, respectively. l1 and l2 denote the half lengths of the pendulums 1 and 2, respectively. Here, M = 1.0 (kg), m1 = 0.3 (kg), m2 = 0.1 (kg), l1 = 0.6 (m), l2 = 0.2 (m), and g = 9"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002863_1.i010002-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002863_1.i010002-Figure4-1.png",
+        "caption": "Fig. 4 Mean distances between underwater agents varied in simulation.",
+        "texts": [
+            " Each computer runs a simulated vehicle along with the necessary MOOS communication processes. D ow nl oa de d by U N IV E R SI T Y O F SA SK A T C H E W A N 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 /1 .I 01 00 02 Simulations were carried out using amaximum of six underwater agents (that is, n 6). Each simulation is run both without the ASVandwith the ASV. To show the performance of theASV, simulations were performedwith varyingmean distances d between theAUVs as shown in Fig. 4. Two particular simulations, called close and spread configurations, are shown in more detail to demonstrate the effectiveness of the ASV. In the close configuration, d 550 m < , where is the communications range parameter; and the AUVs are within communication range of one another. In the spread configuration,d 750 m and communication between adjacent vehicleswas not always possible. Thus, themain role of the ASV in the close configuration is to improve the rate of information propagation in the network"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001162_jssc.200600476-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001162_jssc.200600476-Figure1-1.png",
+        "caption": "Figure 1. Schematic representation of the amperometric detection cell.",
+        "texts": [
+            " The Pt ground electrode from the CE system serves as the counterelectrode providing a sufficiently stable reference voltage [32]. The whole detection system, including the detection cell and electronic circuitry, was placed in a Faraday cage to reduce the background noise. The detection cell consisted of a homemade Perspex container with a volume of 10 mL, filled with the separation electrolyte into which one end of the separation capillary and the Cu-working and Pt electrodes were inserted. A schematic representation of the detector cell configuration is shown in Fig. 1. The working Cu electrode was a 175 lm copper wire sealed into a fused-silica capillary of 180 lm id using a 5-min epoxy glue. For better rugged- i 2007 WILEY-VCH Verlag GmbH & Co. KGaA, Weinheim www.jss-journal.com ness, the fused-silica capillary with the Cu-electrode was then sealed in a glass capillary of 0.5 mm id. The electrode was connected to the measuring circuitry by directly soldering another Cu-wire to the 175 lm copper wire protruding from the other end of the working electrode assembly"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002155_6769-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002155_6769-Figure3-1.png",
+        "caption": "Fig. 3. Model in Lateral Plane with Human in the Loop",
+        "texts": [
+            " It is composed of a torso and two identical legs. Human fixed in the chair can be considered as the torso with 3 degree of freedom (DOF) on the pitchaxis, roll-axis and yaw-axis. Each leg has 6 DOF. As for the control system, 12 DC servo motors are centrally controlled through Baldor NextMovePCI2 by the control computer. The control computer is installed at the back of the waist. The biped robot with human in the loop is modeled with 12-DOF system using the multi-body dynamics method in three dimensional planes as shown in Fig. 2 and Fig. 3. The posture change of human body can be considered as sway around the joint. The motion of biped robots is generated by synchronizing two planar motions. Models in these two planes called sagittal and lateral plane. The origin of the absolute coordinate system coincides with the ankle joint of the supporting leg. To reduce the work of human operator as much as possible, the robot moves automatically by following working program given to the computer in advance. While the robot is moved by the computer, human operator observes the change of the environment and the robot motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003147_1.4007806-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003147_1.4007806-Figure2-1.png",
+        "caption": "Fig. 2 Three kinds of developed faults [22]: (a) spalling, (b) true brinelling, and (c) false brinelling",
+        "texts": [
+            " Many industries replace the defective bearing at its initial stage, however, the goal of the prognostic is to find the amount of time that the machine can survive beyond the time of fault detection, as illustrated in Fig. 1 [3]. Therefore, the progression of the fault in the bearing needs to be monitored. Most of the faults have some kinds of development to failure. One kind of the progression of the defect is that the discrete debris separate from the surrounding area of the first defect since the defect is the place of the stress concentration, which causes the fault to develop as spalling, as shown in Fig. 2(a). Another type of fault progression occurs when the bearing is exposed to excessive load or external vibration that causes true brinelling or false brinelling on the bearing races, as shown in Figs. 2(b) and 2(c) [22], respectively. In these cases, when each rolling element strikes more than one defect in each revolution, it generates impacts which overlap by impulses from other rolling elements. Therefore, the resulting signal is a mixture of different impacts from dissimilar defects. Almost all of the methods in bearing diagnosis try to analyze the signal when an impact is damped in the system before another impact occurs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003045_s11044-012-9326-7-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003045_s11044-012-9326-7-Figure1-1.png",
+        "caption": "Fig. 1 Relations between impact posture and potential barrier",
+        "texts": [
+            " It is not easy to guarantee overcoming the potential barrier only with intuitive control laws in limit cycle walking. It is also difficult to start walking F. Asano ( ) School of Information Science, Japan Advanced Institute of Science and Technology, 1-1 Asahidai, Nomi, Ishikawa 923-1292, Japan e-mail: fasano@jaist.ac.jp from a standing posture smoothly and we must search the suitable initial conditions through a trial-and-error process. The potential barrier in dynamic gait originally comes from the fact that most limit cycle walkers have anterior-posterior symmetric impact posture as shown in Fig. 1(a). A compasslike biped robot [2], for example, must maintain symmetric posture involuntary, and this creates the potential barrier at mid-stance. If the robot can create asymmetric impact posture tilting forwards as shown in Fig. 1(b), this problem can be solved. The easiest way to make the impact posture become asymmetrical is to lengthen the stance leg during stance phases using the prismatic joint while shortening the swing leg. The importance of forward-tilting impact posture has also been discussed in several related works [6\u20138]. On the other hand, the author has wondered about the meaning of anterior-posterior asymmetry for a human foot. As Asano et al. have shown, such as in virtual passive dynamic walking, fully-actuated limit cycle walkers must avoid negative actuator work at every joint for achieving energy-efficient walking [4]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000814_upec.2007.4468943-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000814_upec.2007.4468943-Figure1-1.png",
+        "caption": "Figure 1: stress distribution before modification",
+        "texts": [
+            " Hardened spring steel is one of the materials that have high endurance limit which is more than 400 MPa for 1010 cycles of reverse bending mode hence it has been chosen for this application. To summarise, the flexure spring should have the following specifications:  120mm diameter  Maximum axial displacement of 8mm, half of the stroke.  As high as possible axial stiffness  Less than 0.5mm radial displacement  Maximum stress less than 400 MPa. 4. FE AND EXPERIMENTAL ANALYSIS The spiral spring has six holes at all ends of the three slits to reduce the concentration of the stress around the ends. The slits themselves are the Archimedes spirals. As can be seen from figure 1, higher stresses are still located almost in the outer side of the inner end of the slit and in the inner side of the outer end making stress concentrations around both sides of each slit. Therefore, the spirals have been modified by curling the inner end of the slit into the centre and outer end out of the centre to give a better solution to these stress concentrations. Many simulations have been done to come up with the best curl of the ends. The curl outwards at the outer ends and inwards at the inner ends are shown in fig 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002483_kem.490.97-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002483_kem.490.97-Figure3-1.png",
+        "caption": "Figure 3. Schematic and photograph of Prototype II with top foil removed (see text for symbol definitions).",
+        "texts": [
+            " After a few minutes of operation many of the bump foils detached from the steel support and testing was abandoned. It was concluded that gluing does not provide sufficient bonding strength. As a consequence of the bearing failure, the coating on the top foil was worn through in multiple locations and the exposed hardened steel damaged the surface of the journal as well. Prototype II. Because of the setbacks experienced with Prototype I, it was decided to build a second, significantly redesigned, second prototype. Two important modifications were introduced (Figure 3). The first modification was the new method of attaching bump foils (3) to the half-sleeve (1), which eliminated the need of gluing. In the new design the bump foil sections (3) were mechanically attached to steel blocks (2), which in turn were placed in axial groves machined in the steel half-sleeve (1). The bump foil geometry remained unchanged (Table 1), which allowed the use of the existing hardware in the manufacturing process. However, due to the introduced modification, the locations of bump foil attachment changed for half of the bump foils (compare Figure 2 and Figure 3), resulting most likely in a significantly altered stiffness distribution along the angular direction [11]. The second important change, which was introduced, was replacement of the original steel top foil (coated with low-friction polymeric coating) with a thicker top foil made entirely of a commercially available low-friction polymeric foil (4). The way the top foil was attached to the half-sleeve (1) remained unchanged (5). The new top foil eliminated the danger of damaging the journal if the lubricating film of water is suddenly ruptured and a high-speed rub occurred"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure12.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure12.4-1.png",
+        "caption": "Fig. 12.4-1 A torsional vibration absorber attached to a vibrating system. (Reprinted",
+        "texts": [
+            " In the fi gure, curve 1 corresponds to the value of damping given by Eq. 12.3-36, while curve 2 corresponds to the value of damping given by Eq. 12.3-37, and curve 3 corresponds to the value of damping ratio d = 025. . Next, for constant tuning ( / ), = =d o 1 it can be shown that the optimal damping ratio is given by [12.11] opt = + + + + ( )[ /( )] ( ) 3 1 2 8 1 (12.3-38) VibrationAnalysis_txt.indb 485 11/24/10 11:52:52 AM Principles of Vibration Analysis 486 12.4 Torsional Vibration Absorber Fig. 12.4-1 provides a schematic representation of a torsional vibration absorber [12.14]. It consists of a rotor, with axial moment of inertia Id, attached to a vibrating system, with axial moment of inertia Ip. Th e attachment consists of a torsional damper with damping coeffi cient cd together with a linear torsional spring with stiff ness kd. As with the translational vibration absorber, if the torsional vibration absorber is not properly tuned, the composite system (the original system and the attached absorber) may have a resonance frequency within the frequency range of an exciting moment"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002986_s00170-012-4659-1-Figure16-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002986_s00170-012-4659-1-Figure16-1.png",
+        "caption": "Fig. 16 Schematic of the ECB finishing system for driven tines",
+        "texts": [
+            "0093, while the surface roughness was reduced to 12 \u03bcm and would improve no more than this value even if the roughness of the driving tines was reduced to 2 \u03bcm. According to the experimental results, it could be concluded that improvement of the surface quality of the driving tines make the transmission more exact, but the effect is limited. To improve the transmission precision remarkably, finishing for the driven tines would be needed. 3.2 Finishing process for driven tines The schematic diagram of the ECB finishing system designed for driven tines is shown in Fig. 16. The finishing principle of the driven tines is similar to the generating method. With a layer of insulating cloth covering the machining tool surface, the physical dimension of the machining tool is the same with the driving wheel, which matches this driven wheel. Suppose d1 is the diameter of the driven tine and d is the diameter of the machining tine, the relationship between d and d1 is expressed in Eq. 7: d1 \u00bc d 2h \u00f07\u00de where h is the thickness of the insulating cloth. As shown in Fig. 17, during processing, the relationship between the rotational speed of the machining tool and the driven wheel could be expressed in Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001447_robot.2007.363852-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001447_robot.2007.363852-Figure3-1.png",
+        "caption": "Fig. 3. Torsional Spring and Swivel Damper",
+        "texts": [
+            " vi+1 and vi are the norms of the components of the velocity of point Pi+1 and Pi on the direction ei. 3) Torsional spring: The torsional spring is derived from the angle, a, between two connected segments of the rope. The basic idea is to model each two connected segments as a triangle with a spring as the hypothesis pushing the end points to the full expanded position. The length of the two connected segments remain unchanged. Only the force component orthogonal to the segments is used for the end points (See (a) of Fig. 3). Let e-_- and ei be the unit a ktzPil1 PiH ti11, '+1 = kts i+a 1) fi = -(fi- i + fi+1) -ti+1, (7) (8) (9) where kt, is the torsional spring constant. 4) Torsional damper: The torsional damper works against the torsional spring to prevent any harmonic motion from accumulating. Similar to the linear damper, it also models the internal friction that resists bending in regular objects. Let, vi-1, Vib, be the norms of the velocity components of, vi-1, and, vi, on the direction of, ti_1, and let, vi+1, via, be the norms of the velocity components of, vi+1, and, vi, on the direction of, ti+1, Then, the torsional damper on the points, Pi-1, Pi and Pi+l, can be computed by: ( (Vi-- Vib) + (vi+ -Via) ktdt-1 (10) P P-l-Piu +Pui+1PulI Pui-l-Pi7 (Vi-1-Vib) (vi+ - Via)) ktdti+l (11) (lPi-l Pil lPi+l-Pil )Pi+l-Pil -(fi-i + fi+ ). (12) where ktd is torsional damper constant, vi-1 = vi-1 t i1, Vib = Vi * ti- 1, Vi+ = Vi+ * ti+ , Via = Vi ti+1- 5) Swivel damper: Point, Pi-1, has a velocity relative to the center point, Pi. So far, two components of that relative velocity have been dampened. There still remains a component perpendicular to those two. Without the dampening, point Pi-1 could indefinitely orbit the line formed by extending the edge connecting point P1+j and point Pi (See (b) of Fig. 3). Let s be the unit vector of the swivel dampers of point Pi-1 and P ij, then, -s = e-_ x ei. The swivel dampers can be computed by: f -= ksw (Vi-i-(vi) s HlPi_l -Pill\" (vi+i-vi) * SA i+l PwlPi+i-Pi fi = - (fi- + fi+ ). (13) (14) (15) where k, is the swivel damper constant. vectors with directions from point, Pi-1 to Pi, and from Pi to Pi+j, respectively. Let ti-1 and tij+ be the unit vectors with directions the same as the torsional force applied at the two endpoints and therefore, orthogonal to e-i- and ei respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001109_s11740-008-0080-x-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001109_s11740-008-0080-x-Figure5-1.png",
+        "caption": "Fig. 5 Penetration depth with respect to the modulation frequency and to the powder feed rate",
+        "texts": [
+            " The track generated with the lower powder feed rate have the tendency of a decreasing penetration depth with an increasing modulation frequency. This can be caused by the comparable high local scanning velocity of the laser spot on the surface. Further investigations using a disc laser and WC-Co as filler material have shown similar results, [11]. For all the generated tracks the penetration depth of the cross section was measured from the original surface to the deepest penetration of the keyhole. The penetration depth with respect to the modulation frequency and the powder feed rate is given in Fig. 5. The penetration depth is varying between 0.5 and 2.7 mm. The largest penetration depth was measured at a modulation frequency of 10 Hz and a powder feed rate of 3.6 g/min. For the powder feed rate of 3.6 g/min the clear behavior of the depth with respect to the modulation frequency is shown. Beside the penetration depth of the track the cross sectional area of the tracks was measured. The cross section versus the process parameters is given in Fig. 6. The smallest area is 1.3 mm2 and the largest area is 6 mm2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000296_acc.2006.1657581-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000296_acc.2006.1657581-Figure1-1.png",
+        "caption": "Fig. 1. Schematic diagram of Pendubot.",
+        "texts": [
+            " INTRODUCTION The paper is motivated by applications where the natural operation mode is periodic, orbital stabilization of underactuated systems, enforced by fewer actuators than degrees of freedom, presents a challenging problem (see [3], [11] and references therein). As well known (see [3] and [16]), these systems possess nonholonomic properties, caused by nonintegrable differential constraints, and therefore, they can not be stabilized, even locally, by means of smooth feedback. This paper is devoted to solve a periodic balancing problem for a simple underactuated mechanical manipulator, Pendulum robot (typically abbreviated as Pendubot), whose first link (shoulder) is actuated whereas the second one (elbow) is not actuated (see Fig. 1). Taking advantage on results of Boiko et al. [4], demonstrating that if the combined relative degree of the actuator and the plant is higher than two a periodic motion may occurs in the system with secondorder sliding-mode controllers (SOSM). We apply twisting algorithm ([8]) to drive the Pendubot to a periodic motion centered at the upright position. The twisting algorithm is a recognized tool to generate accurate motions, finitetime stabilization, and robustness properties for mechanical systems ([2], [9], [10])"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002199_6.2009-1852-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002199_6.2009-1852-Figure12-1.png",
+        "caption": "Figure 12. Quad-Rotor vehicle (a) and components (b)",
+        "texts": [
+            " The position reference system allows for modular addition and removal of vehicles, short calibration time, and submillimeter and subdegree accuracy. Data at frame rates of a 100Hz with latency less than 40 ms are obtained for an arbitrary number of vehicles within a specified volume. A sufficient volume is available at the VSTL to enable flight of both rotorcraft and fixed wing aircraft designed for slow flight. Ground vehicles of various types are also supported. V.A.2. Vehicle Description One type of flight vehicle flown extensively within the testbed is a modified remotely-controlled quad-rotor helicopter shown in figure 12a. While commercially available, the onboard electronics are replaced with custom electronics to allow communication with the ground control computers and to enable additional functionality. Other types of air and ground vehicles have also been operated autonomously.55,56 Figure 12b shows an image of the primary vehicle components. A collection of common, modular vehicle hardware and software components have been developed as part of the VSTL facility in order to expedite the vehicle integration process. A custom built vehicle hardware package that includes a microprocessor loaded with common laboratory software, current sensors, voltage sensors and a common laboratory communication system can be applied to a commercially purchased remote controlled vehicle or to a custom vehicle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003738_iros.2011.6094426-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003738_iros.2011.6094426-Figure2-1.png",
+        "caption": "Fig. 2. (a) Photograph of experimental setup and (b) labeled diagram (not to scale).",
+        "texts": [
+            " A custom Matlab script was used to generate actuator control signals which were conditioned through a high-voltage amplifier (Trek Inc.). The actuators were driven open-loop with no feedback. For information on feedback control of a voltage-driven piezoelectric actuator attached to a load, see [13]. Angle of twist was measured in a clamped-free configura tion, using 2D motion tracking software (ProAnalyst, Xcitex Inc.) to calculate ()twist from an edge-on view of the actuator (videos recorded with a PixeLink camera). A photograph and diagram of the experimental setup are shown in Fig. 2. IV. RESULTS A. Fiber orientation For fiber orientation tests, we expect a change in extension-twisting coupling due to changes in the compliance matrix (Eq. 22). These compliance changes give rise to vary ing torque/displacement to voltage relationships as predicted in Sec. 2.1. For a fixed actuator width of W = 2mm, length of L = lOmm, and fiber layer thickness of tc = 80p,m, the orientation of the fiber layer was varied from 0 to 90 degrees. Test actuators were fabricated with 15 degree intervals in orientation (0, 15, 30, 45, 60, 75, 90), and the experimental results are compared to predictions in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000837_6.2008-7192-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000837_6.2008-7192-Figure2-1.png",
+        "caption": "Figure 2. SensorCraft first bending node line.",
+        "texts": [
+            " The primary conclusion is that large amplitude maneuvers at high speed are accomplished using primarily inboard control with the outboard control surfaces being used to reduce the load of the outboard wing. The inboard control surfaces of the SensorCraft configuration are attached closer to the stiff trapezoidal main wing and thus has good control of short period dynamics. The inboard surfaces are close to the first bending node line so that short period control using the most inboard surface helps reduce the energy put into first bending by dynamic control of short period dynamics. The first bending node line is shown in Fig. 2. outboard section to the inboard section. The wind tunnel model was built with 5 control surfaces shown in Fig. 3. It has four trailing edge surfaces, starting just outboard of the break with TE1 to TE4 evenly spaced to the wingtip. There is one leading edge surface at the wingtip labeled LE1. American Institute of Aeronautics and Astronautics 092407 3 TE1 TE2 TE3 TE4 LE1 This paper provides an overview of an optimal control design process for a general SensorCraft configuration as well as the specific design for a wind tunnel test of a SensorCraft model in the NASA Langley TDT wind tunnel"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003229_s11431-012-4986-3-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003229_s11431-012-4986-3-Figure1-1.png",
+        "caption": "Figure 1 Modified FZG gear test rig (a) and schematic diagram (b). Two torque sensors and a temperature sensor were utilized to monitor the working condition of the rig.",
+        "texts": [
+            " In this paper, in order to find an reasonable lubricant quantity which will meet both needs for transmission performance and scuffing resistance for high-speed train gears, the transmission efficiency, bulk and integrate temperature, scuffing load in four different oil immersion depths are investigated by using a modified FZG rig with a potential oil. Then the mean minimum film thickness at the pitch point are calculated by an EHL model in different immersion depths. The ratio of film thickness is used to identify the friction status under scuffing condition. Finally, a reasonable oil amount for gears of high-speed train is suggested based on the relationship of contact pressure and ratio of film thickness at the pitch point. Experiments were completed by a modified FZG back-toback test rig [8] shown in Figure 1. There were two gearboxes, one test gearbox and one slave gearbox. The motor provides the compensation power wasted by the friction of gear transmission loop during the work. To measure the efficiency of test gears, a torque sensor was installed between the output end of the motor shaft and input end of the shaft of the slave gearbox. Meanwhile the clutch to measure the torque in the standard FZG rig was also replaced by a torque sensor. A temperature sensor was located inside the wall of the test gearbox to measure the oil temperature",
+            " The amount of oil for each level, from level 1 to level 4, were 1.25 L, 1.0 L, 0.92 L, and 0.64 L, respectively. The transmission efficiency of the test gears on every load stage, the bulk and integral temperatures, and the load stage when scuffing occurred, were investigated individually. The energy loss of gearbox is mainly caused by the friction of mating gear tooth, churning of the oil and friction of supporting or connecting elements, such as bearings, seals, couplings and clutches. From the experimental rig shown in Figure 1, it can be observed that gear I and gear III are the driving gears while gear II and gear IV are the driven ones. The relationship of torque of these two gear-pairs could be expressed as follows: 4 4 3 34 3 , z T T z    2 2 1 12 1 . z T T z    As, 1 4z z , 2 3z z for FZG test rig, 2 3T T , then, 2 4 4 1 12 34 1 12 34 1 3 , z z T T T z z            1 4 4 12 34 1 1 ,dT T T T            4 IV ,T T d IV IV , ) T T T      \uff08 (1) where Ti ( i=1,2,3,4) stands for the torque of each gear; dT is the compensation torque measured at the end of the motor shaft; IVT is the output torque of testing gear; 12 indicates the efficiency of slave gears"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001078_ecce.2009.5316538-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001078_ecce.2009.5316538-Figure5-1.png",
+        "caption": "Fig. 5: Finite-element flux-plot of the analyzed induction motor",
+        "texts": [
+            " The authors have successfully fitted the model described by (10-12) and (15) to laminated steels typically employed in electric motor manufacturing, such as semi-processed and fully-processed materials under PWM voltage supply [22]. For the estimation of the iron losses in the induction motor a combination of analytical formulations and finite-element analysis is employed. Fig. 4 shows the cross-section of the test induction motor, a 4-pole, 36-slots, 28-bars configuration. The rotor bars are of Boucherot type to ensure high starting torque and high efficiency at rated load. The motor data is given in Annex I. Fig. 5 shows the simulated flux-lines plot for one of the performed experimental test. Unlike in the Epstein test, the typical flux density waveform in electrical machines is nonsinusoidal [17-20]. To account for this, some other authors have extended similar expression for (1) and (2) into the time domain by replacing the frequency with a time derivative [27]. Although not always clearly stated, such an approach is strictly valid only provided that the material coefficients are constant. Also, the slotting effect is the main cause of the minor hysteresis loops and the flux-density waveform especially in the tooth tips may contain reversals, causing minor loops and increased hysteresis loss",
+            " The hysteresis loss and the eddy current loss components are: iron h ec 1 1 N N n n n n P P P = = = +\u2211 \u2211 (18) where: ( ) 22 2 ec,n e n nP k B f n B= (19) ( ) 2 h,n h n nP k B f n B= (20) The computation algorithm of the iron losses is described as follows: 1. A transient magnetic voltage driven 2D finite-element problem is defined. The supply voltage is modelled as sinusoidal 3-phase system, the rotor is kept fixed and the rotor cage is modelled as an open circuit using very high resistivity for the rotor bars and end-rings. In this way, the rotor cage is not shielding the rotor steel any more, as well evident in Fig.5. This analysis was performed using Flux2D software, but the same approach should be taken in any other similar FE package. 2. From the magnetic field solution is extracted the fluxdensity waveform corresponding to an electric cycle for each mesh element from the iron regions, i.e. stator and rotor steel. Note that the shaft region is neglected in this study. The flux-density is analysed per components, radial and tangential. In the induction motor prototype a significant amount of iron losses are produced in the stator yoke, where the flux density has both a radial and a tangential component (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure2.31-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure2.31-1.png",
+        "caption": "Figure 2.31 Specific examples of Nyquist plots",
+        "texts": [
+            " The class \u20181\u2019 system begins with a phase lag of 90 degrees due to the integrator, and again the phase lag increases and the amplitude ratio decreases as before. For stability the locus must pass inside the (\u2212180 degrees, \u22121 0) point on the graph. The class \u20182\u2019 system begins with a phase lag of 180 degrees because of the presence of two integrators in the loop. For this system to be stable some means of \u2018reducing\u2019 the phase lag in the region of the (\u2212180 degrees, \u22121 0) point must be introduced. This is achieved using performance compensation methods that will be described in the next chapter. In order to reinforce the concept of the Nyquist diagram, Figure 2.31 shows Nyquist plots for three specific open loop transfer functions (OLTFs) including the aircraft control system example analyzed in 58 Closing the Loop Section 2.7. The locus for 1/D originates at minus infinity on the y-axis continuing straight up towards the origin as the frequency increases maintaining a constant phase lag of 90 degrees. The second locus of 1/D2 comes from minus infinity on the left towards the origin maintaining a constant phase lag of 180 degrees. In doing so it passes through the instability point (\u2212180 degrees, \u22121) at a frequency of one radian per second",
+            " Doubling the frequency to = 2 0 radians per second shows the open loop gain to be half the magnitude of the = 1 0 radian per second value at \u22126 0 dB which corresponds to Open loop phase (degrees) Alternative Graphical Methods for Response Analysis 65 the closed loop values of \u22127 dB and \u221263 degrees of phase. This is again in agreement with our previous analysis. Let us now go to a more complex example defined by our aircraft control system response generated in Section 2.7 and used in the Nyquist diagrams of Figure 2.31 and 2.32 earlier in this chapter. This same system response is shown on the Nichols chart of Figure 2.37. The most important observation to make here is the clarity of the information presented compared with the equivalent plot of the Nyquist diagram. Because of the logarithmic base, the whole frequency range of interest is visible while showing clearly the important region Open loop phase (degrees) 66 Closing the Loop around the instability point. Again, the ability to derive the closed loop response immediately from the same plot is straightforward and insightful"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000821_icmech.2007.4280058-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000821_icmech.2007.4280058-Figure2-1.png",
+        "caption": "Figure 2 FEM model of stator winding heating",
+        "texts": [],
+        "surrounding_texts": [
+            "I INTRODUCTION\nThe paper is concerned with computational and experimental simulations of stator winding heating of synchronous machine. The synchronous machine operates as high-torque machine with maximal torque 675 Nm at 50 rpm.\nThe aim was to find algorithm for pump control, so that the temperature of stator winding was below safe limit. Software MATLAB/SIMULINK was used for computational simulation of water cooling control. Computational simulations describe direct stator winding cooling by water. Experimental device was used for verification of computational simulations and control algorithm.\nII COMPUTATIONAL MODEL\nThe computational model geometry arises from real synchronous machine. It describes the heat of a part of synchronous machine mainly stator winding. The machine has 36 pair of winding slots and permanent magnets on the rotor. Rotor with magnets is not modelled, because the heat loss is only in the stator winding and rotor effect is negligible on the heating of stator. The brass tubes were comprised in the middle of each winding slots. Cooling water flows in the brass tube. Symmetry of machine was assumed, so only one pair ofwinding slot is modelled. The thermal network method [3] was used for description of machine heating. Thermal networks (Fig. 1) consist from eight nodes. Last three nodes (6, 7 and 8) are used for description of cooling water heating. Thermal model describes transient state, because machine operates with varying load.\nThermal network is possible to be described by differential system equation\nd19i C i+A19 =bi dt\n(1)\nwhere: Ci is thermal capacity concentrated in node i\nMatrix A is matrix thermal conductivities. The matrix A is given by\nA = r12\nA\nwhere:\nA1AS = -,\n1 1 1K1 A2\nJ\\3~As\n-0\nA\nA4\nI\n'1'2 /r,\nAS (2)\nrl + r,L = az01 I+ I+ I+ I r12 r23 r24 I'2\n1 1 _- = (.t403r23O ru 3\nI + I 1 Iu4\n( 02\nA5 I + I + 1\n(2cQ) Q is flow rate c is specific thermal capacity of water rij is thermal resistance between nodes i andj\n1-4244-1 184-X/07/$25.00\u00a92007 IEEE\n2 +K,\nI + 3+K\n1 + 1 24 7'4\n1",
+            "r, is thermal resistance between nodes n and water node r.j is thermal resistance between nodes i and ambient oc is resistance thermal coefficient ZOm is heat losses in node m 4u is ambient temperature\nK1 -(JTwI + K) (r$2 + h-)\n,,2h ( '2-o K13 = (t~2-\nWr + K ) (r2F + 6) (T-3 + hK) __ 2K, 1-S12 =\n(Z'2 + ) (r.3 + 6)\nVector b is given by 01 -A oird u1Irl+K + t'.1 a 'U\n0,r(vo(;rl- S) _ . ~02 F '02I1)u2V02 (7-1 +N)(h 2 02(al3\nb 03- ;+') ),b = {7 K)(r2 K(3+ K) +ru3 03tt\nO .4 r,.,4\n(3) Thermal model was compared with finite elements method model [2]. Parameters of thermal model were corrected so difference between both models was minimal.\nFEM model describe detailed distribution of temperatures and heat flows in individual parts of model. It is more detailed and computational time is too long, so thermal network is more acceptable for control of water cooling in real time.\nIII EXPERIMENTAL DEVICE\nExperimental device (Fig.3) was created for verification of computational simulations. The six thermocouples were placed to each winding slot. The first three thermocouples are above brass tube and other three thermocouples are below brass tube. Infra thermometer was used for the measuring surface temperature.\nThe revolutions of pump motor are controlled by PWM switching converter with power transistor. The right pump capacity is checked by pulse flow meter. Signals from thermocouples, infra thermometer and flow meter are connected with PC via multifunction analog I/0 board (Acquitek). The pump is controlled via multifunctional board too. The user interface was created in software M\\ATLAB (toolbox GUI). The user interface make possible to monitor of individual part device temperatures and set pump capacity. The control algorithm of pump and thermal model is easy added to the interface.\nIV MEASURING\nThe aim of measuring was comparison two system of cooling machine. The first system (common) is the cooling only by means radiation and natural convection. The second system (new) is using direct water cooling in winding slot. The results are summarized in the following table 1.\nThe experimental device thermal phenomena were determined from measuring. Experimental results are used also for correction of thermal model parameters.\nV CONTROL OF WATER COOLING\nThe scheme of water cooling control shows figure 4. Signals from thermocouples are amplified by operational amplifier INA125. Amplified voltages from thermocouples are evaluated by PC to get temperature. The control algorithm calculate flow rate from the temperatures. The value of flow rate is converted to reference voltage used as an input for the PWM switching converter. This process is still repeated.",
+            "Two control algorithms were tested on the experimental device. The first algorithm is on-off regulation and second algorithm is continuous regulation.\nA ON-OFF controller In this case was used on-off controller with switching hysteresis for temperature regulation. If maximal temperature of winding reaches 60\u00b0C then the pump is switched-on. The pump capacity was set on 0.7 1/min. If maximal temperature of winding reaches 55\u00b0C then the pump is switched-off.\n70 1.4\n60\n40\n30ou1020\n0 5 10 15 20 25 30 35 40 Time [nin]\nFigure 5 Measuring results with ON-OFF controller\nThe random load current (heat loss) was generated for testing both continuous controllers. Figure 6 shows simulation results comparison with linear and exponential controller.\n10 1\n0 1.2 '-' 70\nL-\nEE o -\nLi R 30\n1 -\n_0 45\nFigure 5 shows part of ON-OFF regulation. The winding temperature varies between 55-60\u00b0C except first part of graph, where temperature goes up slowly from initial point.\nB Continuous controller In this case the pump capacity depends on the actual winding temperature. The control algorithm has three zones. If maximal temperature of winding is below 40\u00b0C then the pump is switched-off. If maximal temperature of winding is more than 100\u00b0C then the pump is switched on maximal flow rate (0.7 1/min). Two flow rate-temperature relationships were tested between winding temperature 40- 100\u00b0C. These relationships between flow rate and winding temperature are linear and exponential.\n5 10 15 20 25 30 35 40 45 50 55 60\nTime [min]\nSteady state temperature is the same with both controllers for constant load current. The exponential regulation is better than linear, because flow rate is smaller for the same temperature. Other simulations show possibility of long time overloading with exponential controller as shown figure 7.\nVI CONTROL OF WATER COOLING USING HEATING\nTwo ideas are for using thermal predictor model. The first idea is to improve current control system as it is shown in the previous section. The second idea is using thermal predictor model for control of machine heating without thermal sensors in the winding slot.\nA Improving ofcurrent control algorithm The main advantage of prediction is better flow rate control, because winding temperature in the next time is known. Figure 8 shows block diagram of control systems with thermal predictor.\nlo, 11"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003977_j.proeng.2012.09.561-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003977_j.proeng.2012.09.561-Figure6-1.png",
+        "caption": "Fig. 6 The first modal shape in Solid Works",
+        "texts": [],
+        "surrounding_texts": [
+            "The aim of the measurement was to find out natural frequencies and modal shapes of pipeline model for two cases. The first case when springs of elastic support were not preloaded and when they were. Measurement results comparison acquired by Pulse6 for the case when springs were not preload with the results acquired by software Solid Works are presented on next figures (fig.5 - fig.14). Preload of pipeline support springs was created in one half of their length and the second half conserved its previous stiffness. Modal shapes did not differ for both investigated cases but differ in natural frequencies. From this reason we present only natural shapes of the first case, without preload. As we can see from the images there is from fourth modal shape also a deformation of pipeline cross-section and not only distortion of its shape. Corresponding natural frequencies of the particular modal shapes are presented in a table 1. Table 1. Natural frequencies of corresponding modal shapes Natural frequency no. 1. 2. 3. 4. 5. MTC Hammer 20,5Hz 36Hz 82Hz 263Hz 350Hz MTC Hammer P 25Hz 42Hz 85Hz 263Hz 352Hz Solid Works 29,8Hz 35,5 86,3Hz 282Hz 374Hz Where MTC Hammer means the measurement of pipeline vibration by system Pulse when the springs were not preloaded and MTC Hammer P means the measurement of pipeline vibration by system Pulse when the springs were preloaded. The results acquired by software SolidWorks are for the case when springs were not preloaded."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002311_978-3-642-17390-5_19-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002311_978-3-642-17390-5_19-Figure6-1.png",
+        "caption": "Fig. 6. (a) State of agents in configuration 1 at time tp. (b) State of agents in configuration 1 at time tp + 2.",
+        "texts": [
+            " Accordingly, the respective local convergence behaviors are different from each other. Among these, we consider two major configurations that occur frequently: Configuration 1: The peak is located within the convex-hull of the initial positions of agents in the subgroup. The convergence behavior of this configuration can be explained in the following way. For simplicity, we consider a radially symmetric function-profile with a single peak at the center and an initial placement of three agents a, b, and c as shown in Figure 6(a). At time instant tp, agent a remains stationary, agent b makes a deterministic movement toward a, and agent c moves either toward a or b (since, a > b > c). Note from Figure 6(b) that the agent movements at any time instant are within the convex hull of all the current positions of agents. Agent a does not move until after two time steps (i.e., at tp + 2), when b would have crossed the equi-valued contour Ca(tp + 2), leading to the condition b > a. Now, b remains stationary and a starts moving toward b. This cycle repeats, leading to the asymptotic convergence of agents to the peak. Configuration 2: The peak is located outside the convex-hull of initial agent positions and all the agents are situated on one side of the peak"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure8-1.png",
+        "caption": "Fig. 8. Rivet model",
+        "texts": [
+            " 6) was chosen for local I level analysis. Methods for FEM analysis of riveted joints of thin walled aircraft structures 947 The Riveted joint FEM model was built for this region. The Presence of the rivet was taken into account, as well as the distance between middle surfaces of jointed parts. The structure of the model is presented in fig. 7. Dimensions of the model are 400 x 150 mm. Shell elements (Quad4, Tria3) were used. For rivets in the middle, where one can see holes, model presented in fig.8 was used. The model is based on previous work [12,13]. It consists of two circular surfaces, with rivet diameter, connected with rigid MPC element. On the surfaces\u2019 edges, on sheets plans there are GAP contact elements. They model interaction between the rivet shank and sheets. Manufactured head and driven head are not represented. Instead of them displacements in axial direction (z) of nodes of sheets and rivet, on hole edges were tied with MPC elements. Outside the middle part, rivets were modelled as a singular rigid MPC element"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000717_1.2964504-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000717_1.2964504-Figure1-1.png",
+        "caption": "FIGURE 1. Schematic drawing and image of the tribology setup build in the rheometer",
+        "texts": [],
+        "surrounding_texts": [
+            "For testing tribological properties a new device has been designed, which can be mounted as an accessory onto the Physica MCR rheometer. It makes use of the large measurement ranges as well as the motor control mechanism of the rheometer, thus transforming the rheometer into a highly sophisticated tribometer. The setup is based on the ball-on-three-plates-principle (or ball-on-pyramid) consisting of a geometry in which a steel ball is held, an inset where three small plates can be placed, and a bottom stage movable in all directions on which the inset can be fixed. The ball-on-three-plates setup has been used before in a dedicated device to measure static friction coefficients [2]. Figure I depict the new tribometer setup schematically and in photos. The flexibility of the bottom plate is required to get the same normal load acting evenly on all the three contact points of the upper ball. The rotating sphere is adjusted automatically and the forces are evenly distributed on the three friction contacts. An overload of one contact point would result in wrong friction values. The ball as well as the plates for the inset can be exchanged so that the system can be adapted to desired material combinations. The rotational speed appUed to the shaft is producing a sUding speed of the ball with respect to the plates at the contact points. The resulting torque can be correlated with the friction force by employing simple geometric calculations. The normal force of the rheometer is transferred into a normal load acting perpendicular to the bottom plates at the contact points. The rheometer is the core of the setup therefore tests can be run in speed and torque controlled mode. A speed ramp from low to high speeds allows the determination of the friction coefficient against the sUding speed. This plot is also known as Stribeck curve. Stribeck curve measurements performed on the Rheo-Tribometer allow measurements at very low sUding speeds for the investigation of stick slip phenomena and the tribological behavior in the mixed lubrication regime. The determination of the static friction is possible with the same instrument as well when running the tests in torque controlled mode. Similar to the yield point determination in rheology the torque is increase until a movement of the ball can be detected. The system is equipped with Peltier elements in the flexible bottom plate allowing fast and accurate temperature control in the range form -40\u00b0C to 200\u00b0C. In addition a Peltier controlled hood can be used ensuring the desired temperature from the top. The system is not limited on the ball on pyramid test geometry. There is also the possibility to attach a fixture for ball bearings enabUng performance measurements of ball bearings in the full temperature range."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure1-1.png",
+        "caption": "Fig. 1. Reconstructed synchronous precession orbit.",
+        "texts": [
+            " Suppose the signals collected from two coplanar and mutually perpendicular probes are denoted by x and y, respectively. The synchronous frequency component, which is the vibration frequency occurring at 1X rpm and the main consideration in the rotor balancing, in the two directions can be expressed as x \u00bc A sin\u00f0xt \u00feu\u00de y \u00bc B sin\u00f0xt \u00fe w\u00de ; \u00f01\u00de where x is the rotating frequency; u and w are the phase angles; A and B are the amplitudes; t is the sampling time. With Eq. (1) the synchronous precession orbit can be reconstructed as shown in Fig. 1. Generally, the orbit is an ellipse. Besides, since the difference between the Initial Phase Angles of the two directions, u and w, is not always exactly 90 , the coordinates system xoy constructed along with the probes will not be consistent with the major-minor axis gon of the orbit. There is an oblique angle h between the major axis og and the axis ox, see Fig. 1. So neither A nor B in Eq. (1) is the maximum vibration amplitude of the rotor system, balancing analysis and calculation of an anisotropic rotor\u2013bearing system with information collected from some single radial direction is obviously inadequate. In order to further clarify the relationship between mass unbalance and the IPV, here suppose the rotor vibration is excited only by a known mass unbalance, which is placed at angular location aw with a weight of m. The rotor response to this mass then should be x \u00bc mkx sin\u00f0xt \u00fe aw \u00feu0\u00de y \u00bc mky sin\u00f0xt \u00fe aw \u00fe w0\u00de ; \u00f02\u00de where kx and ky are the amplitude amplification factors",
+            " These four parameters change along with the alteration of the working condition related factors, such as rotating speed, load, temperature, probe orientation, etc., but are not influenced by the change of rotor balance state. As long as the working conditions keep the same during the whole balancing procedure, these four parameters are constants no matter in what rotor balance state. The orbit could be further expressed in complex form as r \u00bc x\u00fe jy: \u00f03\u00de A special point, named Initial Phase Point (IPP), shows the location of the shaft center on the orbit in every precession cycle when the key phase event occurs (t = 0), see P0 in Fig. 1. Accordingly, the vector starting from the orbit center and ending at the IPP is named Initial Phase Vector (IPV). For the orbit expressed in Eq. (3), the coordinate of its IPP is P0\u00f0x0; y0\u00de : x0 \u00bc mkx sin\u00f0aw \u00feu0\u00de; y0 \u00bc mky sin\u00f0aw \u00fe w0\u00de: \u00f04\u00de The corresponding IPV is rI \u00bc x0 \u00fe jy0 \u00bc r0 exp\u00f0ja0\u00de; \u00f05\u00de where r0 \u00bc m k2 x \u00fe k2 y k2 x cos 2\u00f0aw \u00feu0\u00de k2 y cos 2\u00f0aw \u00fe w0\u00de h i =2 n o1=2 ; \u00f06\u00de a0 \u00bc arctanf\u00bdky sin\u00f0aw \u00fe w0\u00de =\u00bdkx sin\u00f0aw \u00feu0\u00de g; \u00f07\u00de where r0 is the magnitude of the IPV and a0 denotes the Initial Phase Angle (IPA)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure29-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure29-1.png",
+        "caption": "Fig. 29. Plastic deformations of sheet material a) w1 b) w2",
+        "texts": [
+            " Contact stresses arise on mating surfaces of metal sheets (Fig. 27, 28). During riveting process (using press machine), stresses in rivet hole reach the value of 500 MPa (Fig. 27 \u2013 case w1 and w2). The value of contact stresses after unloading is equal 300 MPa (Fig. 28 \u2013 w1, w2). Jerzi Kaniowski et al. 962 Irreversible plastic deformations of sheet material around the rivet hole remain after the riveting process. The plastic deformation region of sheets after upsetting for two cases of loading is shown in Fig. 29. Decreasing of height of the formed rivet head during riveting process causes the increasing of plastically deformed area around the rivet hole (Fig. 29a, b). Methods for FEM analysis of riveted joints of thin walled aircraft structures 963 Negative stresses (compressive stress state), considerably exceeding yield stress level, appear in the material of sheets around the rivet hole (Fig. 30). Reduced stress distribution during riveting (pressing) is presented in Fig. 31. Jerzi Kaniowski et al. 964 Riveting process has a significant impact on the stress state in joined metal sheets, especially in the case of uniform filling of the rivet hole by the rivet shank"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001059_j.jappmathmech.2009.01.004-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001059_j.jappmathmech.2009.01.004-Figure2-1.png",
+        "caption": "Fig. 2.",
+        "texts": [
+            " 1), three forms of loss of stability can occur: (1) a static loss of stability through a bending form from the action of a compressive axial force P = R2p, since, under the conditions considered, its non-conservative part cannot perform work on the displacements V and W, (2) loss of stability which is also static but occurs through a purely shear mode, and the corresponding critical load is independent of the length of the shell, and (3) dynamic loss of stability which is occurs through a bending-shear mode and can only be revealed by the dynamic method.5 The positions of an element of a shell of length dx are shown in Fig. 2 for the initial undeformed state (a) and for the perturbed state after loss of stability: (b) through a purely shear form, the realization of which is possible under the action of an external pressure p both of a fixed direction as well as of a pressure which remains normal to the surface of the shell, (c) through a purely bending form under the action of a pressure of fixed direction,2 (d) for a dynamic loss of stability which is possible in the case of the action of a follower pressure and only when transverse displacements are taken into account using the Timoshenko kinematic rod model (or any other improved model)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003578_j.cad.2011.03.001-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003578_j.cad.2011.03.001-Figure7-1.png",
+        "caption": "Fig. 7. The singularities and the branch point.",
+        "texts": [
+            " The above observations can be summarized by the following statements: owing to the appearances of two singularities, the selfintersection of the bisector occurs, and the self-intersection point of the bisector is a branch point of the MA; the medial axis disks centered at the points on the self-intersection segment will be outside of the domain, the MA can be obtained by removing the self-intersection segment of the bisector. As mentioned above, the existence of the branch point is accompanied by the singularities of the bisector, which are easy to identify by the conditions given in Eq. (37). Nevertheless, we will give an easier way to identify the singularities in the following. Note that the one-dimensional saddle point algorithm given in Section 5.1 is applicable for calculation of the bisector. As shown in Fig. 7, the calculation starts from p1 and ends at p3 along the direction of the arrow on the bisector. During this process, after reaching the singularity b1 or in the neighborhood, the tangent vector of the next calculation point will reverse with the previous one. However, the point p\u2217 is determined by using the previous tangent vector e1, i.e., p\u2217 = b1+u0e1, thiswill result in two sudden changes of the solutions, i.e., the computation of the tangent point and the value of v. Since the point p\u2217 is far away from the correct segment b1b2, the tangent point on the boundary jumps from q2 to q3 with a sudden jerk"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003689_20121023-3-fr-4025.00027-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003689_20121023-3-fr-4025.00027-Figure5-1.png",
+        "caption": "Fig. 5. MBS model of an electrically driven rear axle of a hybrid vehicle.",
+        "texts": [
+            " (30) Choosing the control input u such that the closed loop dynamics corresponds to (26), the control law reads as u = \u2212k\u03031 |\u03c31| p sgn(\u03c31) + \u03c32 + ( 1 kg + kgJm Jl ) ks\u03b8\u0302 + dm\u03c9m \u2212 dlkgJm Jl \u03c9\u0302l + kgJm\u03b8\u0308ref \u2212 kgJm\u03bb ( \u03c9m kg \u2212 \u03c9\u0302l \u2212 \u03b8\u0307ref ) , \u03c3\u03072 = \u2212k\u03032p |\u03c31| 2p\u22121 sgn(\u03c31), (31) where k\u0303i = kg Jm ki, i = 1, 2. The reference torsion angle is calculated from the reference shaft torque by \u03b8ref = Tref/ks. The sliding mode controller was evaluated with a MBS model of an electrically driven rear axle from a prototype hybrid vehicle of Magna Steyr Fahrzeugtechnik (www.magnasteyr.com). The prototype vehicle is a Fiat Panda with 1200 cm3 combustion engine and an electrically driven rear axle. The maximum power of the electric traction motor is 23 kW, see Schruth et al. (2010). In Fig. 5 the Adams/CarTM MBS model of the electrically driven rear axle is depicted. It shows the motor, the transmission and the differential gear. The three small flat cylinders show the bearing points at the vehicle body. The vehicle mass is considered in the wheel moments of inertia, as the MBS model consists only of the axle. Only 50% of the total vehicle mass are considered in the wheels, since the traction motor of the rear axle supports the combustion engine e.g. with all wheel drive functionality"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003779_s11071-012-0647-0-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003779_s11071-012-0647-0-Figure9-1.png",
+        "caption": "Fig. 9 The ball and beam system. Two variables are the beam\u2019s angle (\u03b8 ), and the ball\u2019s distance (r). Input u is the angular acceleration of the beam",
+        "texts": [
+            " To compare with the novel fuzzy-Pad\u00e9 controller, a pure fuzzy controller, similar to Sect. 3, is also developed. Then we discuss about the performance of the two controllers. It is interesting to see that the fuzzy-Pad\u00e9 controller has superior performance compared to the fuzzy controller. 4.1 System dynamics and the fuzzy controller The ball and beam system is built with two moving parts, a ball and a beam. The ball is free to move along the beam\u2019s direction, and the beam can rotate about the fixed point (Fig. 9). The input to the system is the angular acceleration of the beam. One can control the position of the ball by applying a force to the beam to rotate it. Both the angel of the beam and the position of the ball are measured. Therefore, the model of the system has four states. Those include the position of the ball, r , its velocity, r\u0307 , the angle of the bar, \u03b8 , and its angular velocity, \u03b8\u0307 . The controller uses these variables to derive an error signal used for feedback. The angle of the beam is reduced as the ball nears its desired position, in order to bring the ball to a stop at its desired position with as little overshoot as possible"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001230_12.827720-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001230_12.827720-Figure3-1.png",
+        "caption": "Fig. 3. (a) Schematic of flow cell, (b) photograph of device positioned in flow cell.",
+        "texts": [
+            " For hydrogen peroxide calibration solutions, 10 \u00b5M and 20 \u00b5M solutions were prepared by making successive dilutions from 30% H2O2 stock solution (JT Baker). For glucose calibration solutions, various concentrations between 125 \u00b5M and 20 mM were prepared using D-(+) glucose, 99.5% (Sigma). All solutions were sonicated for at least 10 minutes to equilibrate them with ambient pressure to prevent microbubble formation in the fluidic testbed. In order to test the sensor, it was mounted in a flow cell where the thickness of the sample solution layer is defined by the gasket material placed on top of the sensor (Figure 3). A thin gasket (thickness 100 \u00b5m) was used to mimic the thinlayer conditions of a tear film. An external Ag/AgCl reference electrode (a silver disk electrode with a layer of electrodeposited AgCl) was included in the flow cell. The on-chip working and counter electrodes, as well as the external Ag/AgCl reference electrode were connected to an Epsilon potentiostat (Bioanalytical Systems, West Lafayette, IN, USA) for potential control and signal collection. Proc. of SPIE Vol. 7397 73970K-4 Downloaded From: http://ebooks"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000141_017-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000141_017-Figure1-1.png",
+        "caption": "Figure 1. The adhesive contact between a rigid sphere and an elastic half-space.",
+        "texts": [
+            " By applying the divergence theorem, the force can be integrated over the surfaces instead (Argento et al 1997, Jagota and Argento 1997): F = \u03c11\u03c12 \u222b V2 \u222b V1 \u22072w(s)dV1 dV2 = \u03c11\u03c12 \u222b S2 \u222b S1 n2(G \u00b7 n1)dS1 dS2, where G = x2 \u2212 x1 s3 \u222b \u221e s w(t)t2 dt, F is the force vector, \u03c11, \u03c12 are the number densities of bodies 1 and 2 (number/volume), S1, S2 are the surfaces, V1, V2 are the volumes and x1, x2 are the vectors of the coordinates of a point. Argento et al (1997) proved that the traction on the surface element of body 1 by the surface of body 2 can be written as f = ( \u03c11\u03c12 \u222b S2 n2G dS2 ) \u00b7 n1, where f is the force vector on an element of the surface and n1, n2 are the unit vectors normal to surfaces 1 and 2. The current system is shown in figure 1. If only the vertical direction is considered, the force acting on an element of an elastic half-space by a rigid sphere is p(h) = 4e\u03c3 12\u03c11\u03c12 9 \u222b \u03c0 0 a2 cos \u03b8 sin \u03b8 d\u03b8 \u00d7 \u222b 2\u03c0 \u03b8=0 {h + a \u2212 a cos \u03b8}{[a2 sin2 \u03b8 + r2 \u22122ar sin \u03b8 cos \u03c6 + (h + a \u2212 a cos \u03b8)2]6}\u22121d\u03c6 \u22124e\u03c3 6\u03c11\u03c12 3 \u222b \u03c0 0 a2 cos \u03b8 sin \u03b8 d\u03b8 \u00d7 \u222b 2\u03c0 \u03b8=0 {h + a \u2212 a cos \u03b8}{[a2 sin2 \u03b8 + r2 \u22122ar sin \u03b8 cos \u03c6 + (h + a \u2212 a cos \u03b8)2]3}\u22121d\u03c6, where a is the radius of the sphere, r is the radial coordinate at the surface, h is the distance between the molecules of the sphere and those of the surface and \u03c6, \u03b8are the angles"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000102_156855306776562279-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000102_156855306776562279-Figure3-1.png",
+        "caption": "Figure 3. Model of link i.",
+        "texts": [
+            " The velocities and accelerations of the joints and those of the gravity centers of the links can be derived through time differentiation of (1) and (2). For simplicity, we have set sk = sin(\u03c6k) and ck = cos(\u03c6k) in the formulation. Provided that the snake-like robot only creeps on a plane and behaves as a rope with an infinitesimal diameter, we next formulate the dynamics of the snake-like robot and model the interaction of the robot with the environment. Each link of the snake-like robot on a slope can be modeled as in Fig. 3a. On the basis of the Newton\u2013Euler equation, we can have the following equations for link i: f x i + ff x i \u2212 f x i+1 \u2212 mig sin \u03c8 = mi ix\u0308G f y i + ff y i \u2212 f y i+1 = mi iy\u0308G \u03c4i \u2212 \u03c4i+1 \u2212 ( ff x i si \u2212 ff y i ci ) ( f i \u2212 Gi) + ( f x i si \u2212 f y i ci ) Gi + ( f x i+1si \u2212 f y i+1ci ) ( i \u2212 Gi) = Ii\u03c6\u0308i, (3) where \u03c4i is the torque at joint i, mi and Ii are the mass and the moment inertia of link i, fi (f0 = fn = 0) is the reaction force at joint i, ff i is the friction force at the contact point of link i with the ground, and i = 0, 1, 2, ",
+            " Moreover, since the snake-like robot has no fixed base, from (3) we know that the forces must satisfy: ff + 0f + m0(p\u03080 + g) + m\u03c6\u0308 = 0, (5) D ow nl oa de d by [ U ni ve rs ity o f O ta go ] at 2 0: 16 3 0 D ec em be r 20 14 where: ff = [ n\u22121\u2211 k=0 ff x k n\u22121\u2211 k=0 ff y k ]T \u2208 2, 0f =       \u2212 n\u22121\u2211 k=0 ( mk Gk + k n\u22121\u2211 j=k+1 mj ) ck\u03c6\u0307 2 k \u2212 n\u22121\u2211 k=0 ( mk Gk + k n\u22121\u2211 j=k+1 mj ) sk\u03c6\u0307 2 k       \u2208 2, m0 = n\u22121\u2211 k=0 mk [ 1 0 0 1 ] \u2208 2\u00d72, m = { m(j) } \u2208 2\u00d7n, m(j) =       \u2212 ( mj Gj + j n\u22121\u2211 k=j+1 mk ) sj ( mj Gj + j n\u22121\u2211 k=j+1 mk ) cj       \u2208 2. We have utilized the wheel on our robot to generate the different friction between the tangential and normal directions. The interaction of the robot with the environment is expressed by the Coulomb friction in this study. As shown in Fig. 3b, the friction force along the tangential direction, ff t i , and that along the normal direction, ff n i , are given by: ff t i = \u2212\u00b5tmig cos \u03c8 \u00d7 sign(\u03b4ir t), (6) ff n i = \u2212\u00b5nmig cos \u03c8 \u00d7 sign(\u03b4irn), (7) where \u00b5t and \u00b5n express the friction coefficients in the tangential direction and the normal direction, while \u03b4ir t and \u03b4irn express the displacement at the friction point. The friction forces along the x-axis and y-axis can, therefore, be derived through: ff x i = ff t ici \u2212 ff n i si, ff y i = ff t isi + ff n i ci "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000971_s00170-007-1176-8-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000971_s00170-007-1176-8-Figure3-1.png",
+        "caption": "Fig. 3 Coordinate system of the gear shaving machine with considerations of cutter assembly errors",
+        "texts": [
+            " The crowning mechanism can be further parameterized as shown in Fig. 2, where dv and dh are the vertical and horizontal distances between the pin and pivot at the initial position, respectively. While the pivot (work table) moves horizontally (zt) in shaving from position I to position II, the pin will move a distance dp along the guideway. The rotating angle of the work table y t can be derived as shown in Eq. 1 [10]. The coordinate system of the shaving process can be simplified and illustrated as shown in Fig. 3, where the cutter assembly errors including horizontal, vertical, and centre distance errors are considered. The coordinate systems Ss and S 0 2 are connected to the shaving cutter and the work gear, respectively, while Sd is the fixed coordinate system; S 0 h and S 0 vare auxiliary coordinate systems for importing assembly errors into the horizontal and vertical directions; the angle \u0394h denotes the horizontal assembly error, the angle \u0394v denotes the vertical assembly error, and \u0394E0 indicates the error in the center distance. Other parameters in Fig. 3 are also described as follows: Zt denotes the travelling distance of the shaving cutter along the axial direction of the work gear; C denotes the distance between the pivot and centre of the work gear; \u03b3 denotes the angle between the two crossed axes; E0 represents the centre distance; fs and f2 represent the angles of rotation of the cutter and the gear, respectively, which are related to each other in the shaving operation. y t \u00bc sin 1 dvffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi d2h \u00fe d2v q 0 B@ 1 CA \u00fe sin 1 dh sin q dv cos q \u00fe zt sin qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi d2h \u00fe d2v q 0 B@ 1 CA q \u00f01\u00de If the shaving cutter is assumed to be a helical involute gear, the surface profile and its unit normal can be represented by Eqs. 2 and 3, derived by Litvin [2], where us and vs are surface parameters. rs us; vs\u00f0 \u00de \u00bc xs ys zs\u00bd T \u00f02\u00de ns \u00bc nsx nsy nsz T \u00f03\u00de Tooth profile rs and surface unit normal ns of the shaving cutter represented in the coordinate system Ss can be transformed into the coordinate system of work gear S 0 2 constructed in Fig. 3 by Eq. 4: r 0 2 us; vs;\u03c6s; zt\u00f0 \u00de \u00bc M20s \u03c6s; zt\u00f0 \u00de rs us; vs\u00f0 \u00de \u00bc x 0 2 y 0 2 z 0 2 1 h iT \u00f04\u00de , in which M20s \u00bc M20d Mds Mds \u00bc Mdv 0 Mv 0h0 Mh0e0 Me0f 0 Mf 0i0 Mi0s M20d \u00bc M20a0 Ma0b0 Mb0c 0 Mc 0d There are two major kinematic parameters fs and zt in axial shaving gear with lead crowning so that two meshing equations (Eqs. 5 and 6) are required to calculate the enveloping surface of the work gear, where n 0 2 is derived from ns through the same coordinate transformation mentioned above. f1 us; vs;\u03c6s; zt\u00f0 \u00de \u00bc n 0 2 @ r 0 2 @ \u03c6s \u00bc 0 zt constant\u00f0 \u00de \u00f05\u00de f2 us; vs;\u03c6s; zt\u00f0 \u00de \u00bc n 0 2 @ r 0 2 @ zt \u00bc 0 \u03c6s constant\u00f0 \u00de \u00f06\u00de The parameters fs and f2 denote the angles of rotation of the cutter and gear, respectively, which are related to each other in the shaving operation by Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003155_vppc.2011.6043136-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003155_vppc.2011.6043136-Figure1-1.png",
+        "caption": "Fig. 1. Power split type hybrid system",
+        "texts": [
+            " Using the simulator, a transient characteristics during the mode change were investigated and a control method was proposed to reduce the torque variation the motor torque compensation. The magnitude of the motor torque was calculated from the inertia torque during the mode change. It was found that the torque variation can be effectively reduced by the proposed control algorithm. I. INTRODUCTION A power split type hybrid electric vehicle(HEV) is driven by the internal combustion engine(ICE) and two motorgenerators(MG1 and MG2), which are connected via the power split transmission. The power split transmission has the mechanical path and electrical path as shown in Fig.1. In the mechanical path, the power from the ICE is directly transmitted to the output shaft via the mechanical elements, including the planetary gear. In the electrical path, the battery is charged using the power generated by MG1 (or MG2), while the vehicle is driven by MG2 (or MG1)[1]. In the power split type hybrid electric vehicle, the engine can be efficiently driven using the continuously variable transmission function of the electrically variable transmission (EVT) and two motor-generators (MGs) without any additional power cutoff device[2]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure9-1.png",
+        "caption": "FIGURE 9. Combination of toothed wheels to move a force of 1000 talents with a force of 5 talents (Mechanics 2.21). Drachmann\u2019s rendering of the figure (The Mechanical Technology, p. 82) in Ms L is on the left; the figure from Heronis Alexandrini opera, vol. II, p. 148, is on the right. Neither is exactly faithful to the text, which states that the wheel\u2013axle ratios B:A, D:G\u030c, and Z:H are 5:1, 5:1, and 8:1, respectively.",
+        "texts": [
+            " The wedge and the screw, on the other hand, become more powerful with a decrease in size: the angle of the wedge must be made more acute, and the screw threads more tightly wound. Chapters 21\u201326 of Book 2 discuss the combination of individual wheel and axles, levers, and compound pulleys, each of which is of manageable dimensions, to achieve a large mechanical advantage. As an example we may consider the use of a combination of wheel and axles to move a load of 1000 talents with a force of 5 talents (2.21; Fig. 9). Given the results established in the reduction, this would require a wheel with a radius 200 times that of its axle. But Heron shows that the same mechanical advantage can be achieved by a combination of three wheel and axles with ratios of 5:1, 5:1, and 8:1, respectively. He first describes the construction of a device in which the force exactly balances the weight to be moved, then states that the same force can be made to set the weight in motion by increasing one of the wheel-axle ratios slightly",
+            " Yet a reference to time would be puzzling if Heron had meant to state or imply R\u2032\u2032, since that relationship presupposes that the times of motion are equal for the moving force and the weight (see the next note). And in fact the discussion of the combination of wheel and axles that follows in 2.22 is precisely similar to the discussion of the compound pulley in ch. 24: the comparison is between the times taken by different moving forces applied at the circumference of different wheels to move the weight a given distance (alternatively, it is between the distances they travel in the different times they take to move the weight, assuming they move at the same speed). For example, in Fig. 9 a force of 40 talents applied at the circumference of wheel D must cover 5 times the distance as a force of 200 talents applied at the circumference of wheel B in order to raise the weight by the same distance (the wheel D must turn five times for the wheel B to turn once, and the two wheels have the same circumference). Heron concludes by stating the usual relationship between moving forces: \u201cthe ratio of the moving force (al-quwwat al-muh.arrikat) to the moving force (al-quwwat al-muh. arrikat) is inverse <of the ratio of time to time>\u201d (Opera, vol"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003771_045020-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003771_045020-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of chip holder fabrication process.",
+        "texts": [
+            " In this paper, a method that employs a chip holder with a via hole to spin the photoresist is proposed. The photoresist edge bead can be transferred from the CMOS chip to the chip holder. Indium arrays with 20 m-high, 30 m-diameter bumps are successfully formed on 5 6.5 mm2 CMOS driving circuits for our GaAs/AlGaAs multiple quantum well spatial light modulators. The chip holder with a via hole was fabricated in 4 steps: SiO2 deposition, via hole patterning, SiO2 wet-etching, and inductively coupled plasma (ICP) deep Si etching, as illustrated in Fig. 1. The thickness of the holder should be identical to the CMOS circuit for a smooth surface. First, 3000 nm SiO2 was deposited on the holder wafer in a plasma-enhanced chemical vapor deposition (PECVD) system to serve as a hard mask for the subsequent deep Si etching. Then the wafer was cut into pieces (15 15 mm2/. It is not necessary to have any special mask to pattern the via hole. Reversible AZ5214 was used as negative photoresist, and the CMOS chip itself was placed above the holder wafer as a self-adapting mask"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002565_gt2011-46492-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002565_gt2011-46492-Figure3-1.png",
+        "caption": "Figure 3. Brush Seal Domain: Hybrid Mesh",
+        "texts": [
+            " Domain creation was automated using a parameterised journal file with the six geometric input variables. This approach allowed the careful control of mesh density using both tetrahedral unstructured meshes (within bristle pack) and structured hexahedral meshes (under backing ring), allocated boundary layers, and applied boundary conditions. Higher grid resolution was used in the tip regions and around the last bristle row resulting in a mesh with over 7 million cells. The general approach above used for the initial geometry is illustrated in Fig. 3. A mid-range test pressure of 5 bar total pressure inlet and atmospheric static pressure outlet boundary conditions were defined at the inlet and outlet faces of the domain respectively. Periodic boundaries were attached to the left and right extremities of the domain effectively creating an infinitely repeating linear brush seal. Heat flux boundary conditions applied at the bristle tips and at the rotor-air interface were calculated from experimental measurements of seal power in brush seals carried out at the University of Oxford, assuming that half of the frictional heat generated was conducted up the solid bristles and convected to the air in-between the bristles",
+            " As in the experiment, the fluid temperature at the inlet was set at ambient conditions (300 K). To provide an appropriate heat sink in the initial simulation, and later for the parametric study, the temperatures at the bristle root and backing ring were also set to 300 K. For the purposes of the parametric study, the effect of swirling or tangential flow on the change in the brush seal performance measures is expected to be small as the key geometric inputs are varied. Hence a static wall boundary condition was implemented at the bristle tips for this study. The mesh in Fig. 3 was imported into FLUENT (Version 6.3.26), a commercial code used in the numerical solution of flow and temperature fields. The Navier-Stokes and energy equations were solved using a steady, first order, coupled-implicit finite volume solver. The solution at each iteration was used to update the fluid properties until a converged solution was achieved. Reynolds number between the bristles, based on bristle diameter, was ~300. This showed that viscous forces generally dominate over inertial forces in the brush seal problem"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure5.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure5.1-1.png",
+        "caption": "Figure 5.1 Typical spool valve characteristic",
+        "texts": [
+            " Each of these nonlinearity categories can be treated separately with regard to stability analysis techniques. In the analysis of these discontinuous types of nonlinearity, the issue that we must recognize is that the system response will be dependent not only on the frequency of the stimulus but also on the amplitude of the input to the nonlinearities. Therefore we now have to consider two independent input variables when we evaluate nonlinear systems, first the frequency-dependent aspects of the system and second the amplitude-dependent aspects of the system. Figure 5.1 shows an example of both of the two basic forms of nonlinearity described above. The figure shows the position versus flow characteristic of a typical hydraulic spool valve used to control the flow of hydraulic fluid into and out of a piston actuator. The graph in the figure is typical of a spool valve where the controlling spool land is slightly larger than the ports cut into the sleeve that it rides in. This is called \u2018deadband\u2019 and is favored by the hydraulic system designer as a means to minimize internal leakage of fluid from the pressure side to the return Continuous Nonlinearities 159 side with the attendant unwanted heat generation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002841_j.tsf.2012.04.054-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002841_j.tsf.2012.04.054-Figure7-1.png",
+        "caption": "Fig. 7. Schematic drawing of the formation of a wedge-shaped Co film by placing a shutter between the silicone oil substrate and the metal target during deposition.",
+        "texts": [
+            " In order to quantitatively study the thickness dependence of the wrinkling patterns of the Co film and also to prove the validity of the above measurement technique for film thickness on a liquid substrate, we prepared a wedge-shaped Co film by placing a shutter between the silicone oil substrate and the metal target during deposition. The distance between the shutter and the substrate was fixed to be 1.2 mm. During deposition, the Co atoms can be deposited on the silicone oil surface behind the shutter because of the collision of the gas molecules and other particles. Furthermore, the deposited atoms also can diffuse to the silicone oil surface behind the shutter. As a result, a wedge-shaped Co film naturally forms near the shutter edge, as shown in Fig. 7. The typical wrinkling morphologies of the wedge-shaped Co film are shown in Fig. 8. It is clear that when the distance from the shutter edge x increases, the dimension of the wrinkles decreases substantially. The quantitative relationship between the average wavelength of the wrinkles \u03bb and the distance x is shown in Fig. 9(A). We find that the wavelength \u03bb decreases quickly when xb500 \u03bcm, and then it decreases slowly when the distance x further increases. By using Eq. (5), we estimated the thickness of the wedge-shaped Co film, and the result is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003498_1.4024212-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003498_1.4024212-Figure3-1.png",
+        "caption": "Fig. 3 Dimensions of fluid gap between meshing teeth",
+        "texts": [
+            " Because each tooth has just one rotational degree of freedom and the undeformed flank geometries are known, the lines of flank contact can be determined analytically. As a result, the distances h0i between the theoretical undeformedFig. 1 Gearwheel model with circumferentially moveable teeth 031502-2 / Vol. 135, JULY 2013 Transactions of the ASME Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use flanks can be calculated. They can be positive or negative depending on whether the flanks are in contact or not, cf. Fig. 3. The teeth are connected to their respective wheel body by spring-damper elements. The stiffnesses ci of the springs between teeth and wheel body are determined according to the ISO 6336 standard [37]. For a tooth pair at the pitch point C the stiffness related to face width is cpitch \u00bc 14N=\u00f0mmlm\u00de which includes tooth bending, shear deformation, and Hertzian deflection. This value was verified by static FE simulations. The stiffness is assumed to decline quadratically with the distance g between the actual mesh point and the pitch point because the lever arm of tooth bending raises [35,38]: c g\u00f0 \u00de \u00bc cpitchCR 1 1 SR\u00f0 \u00de g gmax 2 \" # (1) with the gear blank factor CR (contribution of the gear wheel body to the overall contact stiffness), the ratio of maximum to minimum stiffness SR, and the maximum distance between possible mesh points and the pitch point gmax",
+            " The integrator scheme determines how often and at which points in time the force element is executed and if the integration process is iterative or not. 2.2 Analytical Gear Force Calculation. The basis for the analytical approach is the one-dimensional Reynolds equation for thin fluid films [10] including wedge and squeeze terms @ @xg qh3 g @p @xg \u00bc 6 @ @xg qhU\u00f0 \u00de \u00fe 2 @ qh\u00f0 \u00de @t (2) with the fluid pressure p, the oil density q, the oil viscosity g, the gap height h, the time t, and the sum velocity in the xg direction U. The coordinates as well as the dimensions of the fluid gap are shown in Fig. 3. The pressure gradient in the yg direction as well as the velocity in the yg direction may be ignored, because the dimension of the flanks in the yg direction is much longer than in the xg direction and the velocities are fairly small. The squeeze term @ qh\u00f0 \u00de=@t represents transient gap effects and might be interpreted like a normal damping term [39]. The undeformed gap height is assumed to be parabolic and you get (cf. Fig. 3) h \u00bc hmin \u00fe hs xg \u00bc h0 \u00fe hg\u00f0xg\u00de \u00fe w\u00f0xg\u00de \u00bc h0 \u00fe x2 g 2R \u00fe w xg (3) with the elastic deformation w and the combined flank radius of curvature R. Because the fluid gap is treated one dimensionally, flank corrections in the transverse yg direction, like crowning, cannot be investigated. Nevertheless, profile corrections in the xg direction like tip relief can be regarded. The Reynolds equation (2) is solved differently depending on the load. For small forces the elastic flank deformation w can be ignored and the fluid may be treated as isoviscous; therefore, the Reynolds equation can be integrated directly",
+            " This problem can be solved completely analytically but the resulting equations are quite long and therefore not given here. The solution is valid for small forces and big gap heights. Therefore dry and mixed friction may be ignored at this point. 2.2.2. Elastohydrodynamic Solution for High Flank Forces. At high forces the viscosity depends on pressure and the tooth flanks are deformed. To solve the problem, the shape of the gap is assumed to be equal to a dry Hertzian contact. The basis of this theory is an elastic half-space. In the middle of the contact the gap height is constant (parallel gap), cf. Fig. 3. The parallel contact area has the size 2a 2b, with b\u00bc L/2 and the length of the contact line L. The contact footprint half-width is given by [40] a \u00bc ffiffiffiffiffiffiffiffiffiffi 2RFn pbE0 r (4) with the normal force Fn and the reduced Young\u2019s modulus E0, Journal of Tribology JULY 2013, Vol. 135 / 031502-3 Downloaded From: http://tribology.asmedigitalcollection.asme.org/ on 01/28/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 1=E0 \u00bc 1 2 1 E1 \u00fe 1 2 2 E2 (5) The Young\u2019s moduli of the gears are E1 and E2 and 1;2 are the Poisson\u2019s ratios",
+            " Dry friction, pressure dependency of viscosity and density, as well as tooth flank deformations are taken into account. An approximation of the bodies by semi-infinite half spaces is applied for the description of the elastic deformation. The flanks are discretized and the gradients in the Reynolds equation are approximated by finite differences. Reynolds conditions (also known as Swift Steiber conditions) are used as boundary conditions, which means that p\u00bc 0 and @p=@xg \u00bc 0, respectively, @p=@yg \u00bc 0 at the edges of the pressure field [10]. The undeformed gap height hg(xg, yg) (cf. Fig. 3) is calculated for the real involute flank profile at each grid point. Therefore arbitrary flank corrections like tip relief or transverse crowning can be regarded. The problem is solved using a multilevel multi-integration method according to Venner and Lubrecht [42]. The algorithm was extended by transient terms as well as mixed friction terms. The transient terms were included using a second order backward discretization. For this purpose, the gap forms of each tooth contact need to be saved for the last two time steps"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure4.2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure4.2-1.png",
+        "caption": "Fig. 4.2-1 Synchronous whirl of a shaft due to mass unbalance.",
+        "texts": [
+            " These forces produce an unbalance moment with a torque a rm 2. If we add a balance mass (1/3)m on each wheel, the two balance masses will produce a moment with the same magnitude but opposite direction to the original unbalance moment. Then the crankshaft will be balanced. Fig. 4.1-5 depicts this balancing solution. 4.2 Whirling of Rotating Shaft To establish some important concepts and defi nitions and to provide information of general utility, it is useful to consider the idealized system represented in Fig. 4.2-1 [4.1, 4.2, 4.3]. Th e system consists of a disk of mass m located on a light shaft supported by two bearings. Th e mass center G of the disk is slightly off the shaft axis. In this model, an infi nitely rigid simple support is assumed for the bearings. Th e rotating shaft tends to bob out at certain speeds, and whirl takes place in a complicated manner. Whirling is defi ned as the rotation of the plane made by the bent shaft and the line of centers of the bearings. Th e phenomenon results from various causes such as mass unbalance, shaft fl exibility, hysteresis Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001930_je900770b-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001930_je900770b-Figure5-1.png",
+        "caption": "Figure 5. Cross section of the T-cell used to perform chronoamperometry.",
+        "texts": [
+            " The platinum microdisc working electrode was polished on soft lapping pads (Kemet Ltd., U.K.) using alumina powder (Buehler, IL) of (5.0, 1.0, and 0.3) \u00b5m sizes. The electrode diameter was calibrated electrochemically by analyzing the steady-state voltammetry of a 2 mM solution of ferrocene in acetonitrile containing 0.1 M TBAP, with a diffusion coefficient for ferrocene of 2.3 \u00b710-5 cm2 \u00b7 s-1 at 293 K.24 The electrodes were housed in a glass cell \u201cT-cell\u201d designed for investigating microsamples of ionic liquids under a controlled atmosphere25 (Figure 5). RTILs are sensitive to water,25,26 and the presence of water can alter the viscosity of the ionic liquid and reduce the electrochemical window; therefore, the samples are purged under vacuum before voltammetry is carried out. In addition, this procedure removes electroactive oxygen from solution.27 The working electrode was modified with a section of disposable micropipet tip to create a small cavity above the disk into which a drop (20 \u00b5L) of ionic liquid was placed. DMPD was directly dissolved in [C4mim][BF4] at concentration of 20 mM"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001022_j.jsv.2008.05.023-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001022_j.jsv.2008.05.023-Figure1-1.png",
+        "caption": "Fig. 1. Structure of an electric disk grinder.",
+        "texts": [
+            " And a new gear-element is developed, in order to describe the nonlinear contact-impact behavior of bevel gears, which use relative displacement\u2013force pattern to solve dynamic behavior of bevel geared rotor system efficiently. Moreover, an analysis program, GEARS, is developed to simulate the nonlinear dynamic behavior of geared rotor system since power switch on. ARTICLE IN PRESS Q. Gao et al. / Journal of Sound and Vibration 319 (2009) 463\u2013475 465 An electric disk grinder consists of the following main components: an armature shaft, housing and bearings, a pair of bevel gears, a brush, a fan, and a grindstone (Fig. 1). A dynamics model, which includes the above components, is developed. The motor torque Tm, which is generated by armature, and depends on the time and rotational speed of the armature, is given by Tm \u00bc T0\u00f01 cos\u00f04pf mt\u00de\u00de (1) where the generating power of motor T0 is a function of the rotational speed of motor, fm is the frequency of the motor alternator, and t is the time. The gear, assumed to be a rigid disk with six degrees of freedom (axial and lateral displacements, and axial and lateral rotations)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure10-1.png",
+        "caption": "Fig. 10 The linkage subsystem: a simple, closed kinematic chain of eight rigid links and eight joints",
+        "texts": [
+            " Assume that an inertially axisymmetric rigid body of a multibody system moves so that its angular velocity is composed of: (a) a spin about its axis of symmetry S and (b) a precession about an axis 2 \u2013 normal to S \u2013 fixed to an inertial frame, parallel to one of its transverse principal axes of inertia, about its centre of mass, and offset with respect to S. Then, the contribution of the body to the C-matrix of the system vanishes. Apparently, the horizontal bevel gears 9 of the RGB comply with the assumptions of Theorem 1, and hence the contribution C\u0304RGB of the two RGBs to the C-matrix of the system vanishes, i.e. CRGB~O44 \u00f058\u00de As pointed out in section 1, the linkage is composed of eight rigid links coupled by eight massless joints, which form a simple,* four-DOF closed kinematic chain, as depicted in Fig. 10. In the figure, the robot frame (not shown) is the eighth link; it is common practice in linkage kinematics to number the linkage frame as 0, which is done here. All moving links are thus numbered from 1 to 7. No confusion should arise between these link numbers and the numbers of the EGT and the RGB, as the numbering of each subsystem is confined to its own subsection or subsubsection. Furthermore, all links bear a mnemonic name, besides their number: 2I, 2II are the panning discs; SI, SII are the shoulder brackets; WI, WII are the * A kinematic chain is said to be simple when none of its links is of an order higher than binary, i",
+            "comDownloaded from wrist brackets; and M is the moving platform. Henceforth, any link property will be indexed by its link number, with the provision that, with the exception of the seventh link, all odd-numbered links belong to limb I and their even-numbered counterparts to limb II. Next, the twist-shaping matrices of all the moving links of the LK subsystem are obtained. These matrices map motor rates into LK link twists. It should first be noticed that the contributions of the panning discs 1 and 2 of the linkage (Fig. 10) have already been taken into account within those of the planet carrier 8 of the EGT subsubsystem, as each panning disc is rigidly attached to its corresponding planet carrier. Hence, the contributions of these links are not accounted for in this subsubsection. Let t3, \u2026, t7 be the twists of bodies SI, SII, WI, WII, and the MP respectively. Expressions for the twist arrays and the twist-shaping matrices for limb I are derived below, the counterpart expressions for limb II being similar but associated with even-numbered links. Keeping the number labelling of Fig. 10, the shoulder bracket SI, whose centre of mass is assumed to coincide with its centroid, is first considered, so that its twist t3 and its twist-shaping matrix T\u03043 are t3~ _hI1 _c3 \" # : 1 0 k|aI2 f I|aI2 \" # |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl} S3 [ R4|2 _hI : T3 _q, T3:S3KIE [ R4|4 \u00f059\u00de where KI was defined in equation (21). It should be noticed that, for limb II, matrix KII should replace KI. For the wrist brackets WI, recalling the property of the velocity field of a rigid body under Scho\u0308nflies motion and equations (18a) and (24b), the expressions below for the twist array t5 and the twist shaping matrix T\u03045 are derived t5~ _hI1 _pI \" # : eT1 SIA {B \" # |fflfflfflfflfflffl{zfflfflfflfflfflffl} S5 [ R4|4 _h: T5 _q, T5:S5E [ R4|4 \u00f060\u00de Finally, the twist array t7 of the MP is equal to the twist array t of the MP at the operation point P, which is given in terms of the vector of the actuated joint rates h\u0307 in equation (16), whence t7:t~A{B _h~A{BE|fflffl{zfflffl} T7 _q [ T7~A{BE \u00f061\u00de Further _ T3~ _S3KIE, _S3~ 0 0 k| _aI2 _fI|aI2zf I| _aI2 \u00f062\u00de while, from equation (60) _ T5~ _S5E, _S5~ 0T 4 _SIA {BzSI A{ : BzSIA { _B \" # \u00f063\u00de where 04 is the four-dimensional zero vector, while S\u0307I and (A{)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003382_978-1-84996-062-5_5-Figure4.37-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003382_978-1-84996-062-5_5-Figure4.37-1.png",
+        "caption": "Figure 4.37 The principles of the Globo technique for welding plastics",
+        "texts": [
+            " The weld requires firm clamping such as 400N for 3mm thick plastic [52]. This need for clamping is possibly the Achilles heal of the process since if one is clamping then it is as easy to use heated bars to make the weld, but possibly this is not so quick. However, the clamping for laser welding can be done by using a transparent plastic or glass plate to press on the stack (as shown in Figures 4.35, 4.36) or a transmissive ball through which the laser passes and is focused; the ball rolls along the required seam as shown in Figure 4.37 \u2013 this technique is known as Globo welding. With the glass plate method and a powerful enough laser, the beam can be shaped by optics (e.g., ring optics [53]) or masks to allow a complete circular weld or other shape to be done in one pulse. Alternatively the weld could be made in a quasi-simultaneous manner by scanning in the required pattern. In transmission welding of plastics: \u2022 the welds are strong, watertight and without any surface disruption, since the melt zone is only at the interface with the absorber; \u2022 it is not necessary to have direct contact with the welding tool; \u2022 the process is fast; \u2022 there is greater design flexibility; and \u2022 it is a clean process with no fume of dust"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.4-1.png",
+        "caption": "Fig. 3.4. Spherical wrist at a singularity",
+        "texts": [
+            " Notice, however, that this form of Jacobian does not provide the relationship between the joint velocities and the end-effector velocity, but it leads to simplifying singularity computation. Below the two types of singularities are analyzed in detail. On the basis of the above singularity decoupling, wrist singularities can be determined by inspecting the block J22 in (3.43). It can be recognized that the wrist is at a singular configuration whenever the unit vectors z3, z4, z5 are linearly dependent. The wrist kinematic structure reveals that a singularity occurs when z3 and z5 are aligned, i.e., whenever \u03d15 = 0 \u03d15 = \u03c0. Taking into consideration only the first configuration (Fig. 3.4), the loss of mobility is caused by the fact that rotations of equal magnitude about opposite directions on \u03d14 and \u03d16 do not produce any end-effector rotation. Further, the wrist is not allowed to rotate about the axis orthogonal to z4 and z3, (see point a) above). This singularity is naturally described in the joint space and can be encountered anywhere inside the manipulator reachable workspace; as a consequence, special care is to be taken in programming an end-effector motion. Arm singularities are characteristic of a specific manipulator structure; to illustrate their determination, consider the anthropomorphic arm (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001470_j.jsv.2008.05.001-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001470_j.jsv.2008.05.001-Figure5-1.png",
+        "caption": "Fig. 5. Phase-plane plots of the pendulum with varying coefficient of restitution (13 cycles). Initial condition: ( 1, 1,0,0). (a) y1 (w \u00bc 0.9), (b) y2 (w \u00bc 0.9), (c) y1 (w \u00bc 0.8), (d) y2 (w \u00bc 0.8), (e) y1 (w \u00bc 0.7), (f) y2 (w \u00bc 0.7), (g) y1 (w \u00bc 0.6) and (h) y2 (w \u00bc 0.6).",
+        "texts": [
+            " Hence, the value of k is kept constant for all the simulations and the coefficient of restitution is varied through the damping coefficient c. 3.1. Inelastic impacts (w40) The following values of the system parameters were used for the simulation: m1 \u00bc m2 \u00bc 0:1 kg; l1 \u00bc l2 \u00bc 0:1m; k \u00bc 1012 Nm; c has been varied For w \u00bc 1, the energy of the system is conserved (Fig. 4). The response of the system is aperiodic but follows a repeated pattern. There are 5 impacts per cycle (refer to Fig. 4(c) and (d)). In the phase-plane plot, with 0owo1, a gradually collapsing set of curves is obtained. As w is decreased (Fig. 5), the behaviour becomes more ordered, the number of impacts per cycle decreases and the variance in ARTICLE IN PRESS S. Singh et al. / Journal of Sound and Vibration 318 (2008) 1180\u201311961186 angles of impact over subsequent cycles reduces. In all cases, on the first impact, the pendulum jumps into a gradually collapsing \u2018attractor\u2019 after the first impact itself. The peripheral trajectory of this attractor can be termed as the limiting curve. There is a fixed-energy limiting curve (phase space) of the system, independent of the initial conditions and the coefficient of restitution. If the system starts from a point (phase space) inside the limiting curve, then it traverses the characteristic trajectory in the gradually collapsing attractor (see Fig. 5). But if the system starts from a point outside the limiting curve then it quickly jumps inside the curve after some initial impacts. We observed that while starting from outside the limiting curve, the system may not necessarily jump on the peripheral trajectory of the attractor (i.e., the limiting curve) but start on a lower trajectory close to the peripheral one. Note that for 0.7owo1 (Fig. 5), there are five impacts in every cycle. For w \u00bc 0.6, there are four distinct impacts. For values of wp0.6 there are a large number of small impacts till the two links separate completely when both link angles are 0 rad, that is, when both links are vertical. As angular momentum (about the upper hinge) is conserved during impacts, we expect that there be no jumps in the curve (see Fig. 6). For wo1, the angular momentum varies sinusoidally with the amplitude of ARTICLE IN PRESS S. Singh et al. / Journal of Sound and Vibration 318 (2008) 1180\u20131196 1187 successive peaks decaying exponentially (see Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002213_tdc-la.2010.5762870-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002213_tdc-la.2010.5762870-Figure1-1.png",
+        "caption": "Fig. 1. Transmission tower characteristics",
+        "texts": [
+            " The outline of the paper is as follows: Section II describes the main features of the tower type and also the transmission line used for the study. In Section III, the main simulation results related to the research topic are presented. Finally, in Section IV the main conclusions drawn from the study are presented. A 978-1-4577-0487-1/10/$26.00 \u00a92010 IEEE The phase conductors of the considered transmission line are horizontally arranged each having a quadruple bundle configuration (4 x Rail), as depicted in Fig. 1. The transmission line considered during the study is located in the southern region of Brazil and operates at 525 kV. The shield wire within the segment chosen corresponds to a CAA Dotterel wire. The soil resistivity and tower foot resistance considered are 1000 m and 20 , respectively. The line segment where the electromagnetic induction analysis is carried out was represented by PI parameters (LCC \u2013 Line/Cable constants) with 0.5 km length each (Fig. 2). It should be mentioned one constraint encountered with the shield wire along the line",
+            " The reason for this preference is because some preliminary tests, as it will be shown later, showed that the contribution of the latter type of induction is smaller when compared to the former effect. Figure 2 [11], shows the two type of induction effects present in a Transmission line. It can be seen that the electromagnetic induction has to do mainly with the current present in the phases (variation of the electromagnetic field around the conductor). Thus, inducing both voltages and currents in another conductor (shield wire), whereas the electrostatic effect is related to the coupling effect (capacitance) between phases and also with the ground. As for the tower configuration showed in Fig. 1, it was of our interest to first determine what was the best shield wire circuit configuration producing the highest tap-off power. Thus, two configurations were considered: i) Use of only 1 shield wire (or both working independently) with the earth as a return path. ii) Use of both shield wires in a closed-loop configuration so So, it was adopted the closed-loop configuration depicted in Fig. 3(b), as in this case the magnitude of the induced current nearly doubled. In Fig. 3(b), Rg1 refers to the bypass conductor used to link both shield wires (SW1 and SW2). Rg2 and Rg3 are the equivalent tower foot resistances, whereas Rg4 refers to the equivalent small load connected to the closed-loop configuration. From the analysis and simulations conducted, the following results have been obtained. Regarding the tower configuration shown in Fig. 1, it was observed that the induced current in one shield wire, say SW2, is shifted 180\u00b0 of the other shield wire. This is advantageous because by connecting in series both wires (SW1 and SW2) the equivalent current will have nearly twice the value of the magnitude when only one wire would be operating independently. For some other tower configurations (say vertical configuration of the phase conductors) this value will be different. Figures 4(a) and (b) show the induced current and the power dissipated on the load resistance (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002191_6.2009-1803-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002191_6.2009-1803-Figure4-1.png",
+        "caption": "Figure 4. Aircraft coordinate system and potential C.G. shift.",
+        "texts": [
+            "16 Our design goal is to then select a set of parameters such that all estimated quantities are slowly varying and all system dynamics may be adequately captured. Given the above, we now present the parameterization chosen for this work. This primarily entails identifying the appropriate regressor vector. In Reference 2, analysis was performed on the change in the equations of motion for a damaged aircraft where the aircraft C.G. has shifted due to some loss to the structure, as illustrated in Figure 4. Through determination of the angular momentum for the system shown, we find that Equation 52 specifies the aerodynamic moments for the damaged aircraft. American Institute of Aeronautics and Astronautics 13 \u20ac Maero = L M N           = I d\u03c9 dt +\u03c9 \u00d7 I\u03c9 + \u0394r \u00d7 Faero (52) Here, \u0394r is the shift in the C.G. location, \u03c9 is the angular velocity of the body-fixed frame (a vector [p q r]), I is the inertia tensor (taken about the new C.G. location), and Faero represents the aerodynamic forces. The aerodynamic forces on the system are then represented as the following: \u20ac Faero = X Y Z           = m \u02d9 V + m d\u03c9 dt \u00d7 \u0394r + m\u03c9 \u00d7 (\u03c9 \u00d7 \u0394r) + m\u03c9 \u00d7V \u2212G (53) Here V is the aircraft velocity as represented in the body-fixed frame and the over dots denote the relative linear accelerations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002673_02533839.2012.751334-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002673_02533839.2012.751334-Figure1-1.png",
+        "caption": "Figure 1. Phase paths of the second-order system.",
+        "texts": [
+            " Hence, j _r(t1)j j _r(t0)j \u00bc q\\1: (18) Extending the trajectory into the half plane, r(t)\\0 after similar reasoning achieves successive crossing of the axis r(t) \u00bc 0 to satisfy the inequality j _r(ti\u00fe1)j= j _r(ti)j \u00bc q\\1. Therefore, its solutions cross the axis r(t) \u00bc 0 from quadrant II to quadrant I, and from quadrant IV to quadrant III. Every trajectory of the system crosses the axis r(t) \u00bc 0 in finite time. After gluing these paths along the line r(t) \u00bc 0, we obtain the phase portrait of the system, as shown in Figure 1. This algorithm features a twisting of the phase portrait around the origin and an infinite number encircling the origin occurs. According to Levant\u2019s papers (1993, 2003), the total convergence time is estimated as T X j _r(ti)j j _r(t0)j(1\u00fe q\u00fe q2 \u00fe ) \u00bc j _r(t0)j 1 q : (19) As a result, we can show that the trajectories perform rotations around the origin while converging in finite time to the origin of the phase plane. The finite time convergence to the origin is due to switching between two different control amplitudes as the trajectory comes nearer to the origin (Khan et al"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002157_icctd.2009.202-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002157_icctd.2009.202-Figure6-1.png",
+        "caption": "Figure 6. The r radius",
+        "texts": [
+            "The virtual wave hill In this formula, R is maximum radius of the wave and r is radius of cell from center. The virtual wave hill with two radiuses 3 and 5 has shown in fig 5. Wave value is center is high and as we go forward to outside of wave, values become low. This wave causes to the robot escapes from local minimum and moves to the goal. 3. Virtual Target In this method, with using a virtual target in around suddenly obstacle, we try to pass from suddenly obstacle and escape from local minimum. At first, for determining the virtual target's distance, we need to r radius. The Fig 6 shows r radius. The size of r depends on obstacle's magnitude. With using this radius, we gain one point around the robot. We select this point over the sudden obstacle. Nearest point to basic target with distance r as virtual target's point, that it hasn't obstacle's cell. Fig 7 shows point and virtual target's point. Than without regarding first wave, we expand a new wave from virtual target with using first wave algorithm. this wave continues until reaching robot point and in case of reaching robot point, it stops and the robot moves in base of this new wave to virtual target ,it passes from suddenly obstacle and escapes from local minimum, then it moves according to first wave to basic target"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002088_cdc.2010.5717137-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002088_cdc.2010.5717137-Figure1-1.png",
+        "caption": "Fig. 1. Robot geometry and reference frames.",
+        "texts": [
+            " The ith robot configuration, qi, specifies the position and orientation of a moving Cartesian frame FAi , embedded in Ai, with respect to a fixed Cartesian frame FW . It is assumed that the robots\u2019 paths can be planned in concert, based on a set of given and fixed initial and final goal configurations, Q = {q0i ,qfi | i \u2208 IA}, where IA is the set of unique identifiers representing the N robots. In this problem formulation, the robot state or configura- tion is defined as qi = [xi yi \u03b8i] T \u2208 R 3, where (xi, yi) and \u03b8i are the coordinates and orientation of FAi with respect to FW , respectively. The robot\u2019s rotation \u03b8, defined in Fig. 1, is restricted to the range \u0398 = [\u03b8min, \u03b8max]. Without loss of generality, the method is presented for C = W\u00d7\u0398, where W \u2282 R 2. It is assumed that minimal distance can be represented by a convex cost function modeled by the square of the L2-norm, J = \u222b T 0 N \u2211 i=1 (qi(t)\u2212 qi(t\u2212 1)) 2 (1) Based on the above assumptions, the coordinated geometric path planning problem can be formulated as a mixed-integer linear program (MILP) that can be readily solved using the TOMLAB/CPLEX software, using the approach presented in the following sections"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001334_iembs.2009.5333377-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001334_iembs.2009.5333377-Figure2-1.png",
+        "caption": "Fig. 2. The spirometer and pulse oximeter sensors to be interfaced with the handheld device.",
+        "texts": [
+            " This examination of needs has identified that (a) reduction in cost of the medical device and (b) educating the patient on the usage of the device, are two ways to approach this problem. The reduction in cost will help make healthcare affordable to a larger population, primarily consisting of lower and middle \u2013 income groups afflicted with respiratory and lung diseases. Figure 1 shows the concept model of a handheld device being designed with the ability to be connected to a number of different modules which would assist medical practitioners in diagnosing a patient. Figure 2 shows the spirometer and pulse oximeter sensors to be interfaced with the handheld device. In addition a temperature sensor will also be included. Along with these modules a Universal Serial Bus (USB) connection will provide increased usability for the end user if a personal computer is available. A. Low-Cost Alternatives As summarized in table 1, cost is still a major consideration for medical devices in India. As the medical device companies in developed countries spend hugely on improvement of their existing products, it increases the cost of healthcare"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002656_20110828-6-it-1002.02550-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002656_20110828-6-it-1002.02550-Figure5-1.png",
+        "caption": "Fig. 5. The kinematic scheme of the Motoman IA20 manipulator",
+        "texts": [
+            " Regarding the motor velocities, there is anyway a security threshold q\u0307max. If a single motor passes the maximum allowed velocity value, the whole vector q\u0307 is reduced by multiplying it by this ratio: q\u0307max max(|q\u0307|) . Using this solution, we do not change the end effector velocity direction, but only the intensity, thus we do not change the path. We have tested the control system on a 7 DOFs spatial manipulator arm, with the same kinematic structure of the Motoman IA20 robot. This manipulator, as shown in Figure 5, can be seen like two spherical robots, the first three joints and the last three joints (spherical wrist), coupled with a joint located between the two sphere centers, the fourth one. Finally we have tested the control system on a 7 DOFs planar manipulator, in an environment cluttered with several obstacles. The first experiment has been the test of the path planner making the manipulator pass between two ellipsoidal obstacles. Figure 6 shows the capability of the control system to avoid collisions and pass inside the corridor between the two obstacles"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001559_acemp.2007.4510568-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001559_acemp.2007.4510568-Figure2-1.png",
+        "caption": "Fig. 2. Studied portion",
+        "texts": [
+            " The two double excitation winding currents have the same values and opposite directions [2]. The developed model is based on the main flux path and a part of the magnets leakage flux\u2019s path. In this paper, a 3D magnetic equivalent circuit is proposed and its validity is established by comparing the results it gave to those given by a finite-element model. In addition to the polar symmetry, the homopolar structure has a plan of symmetry. This plan is noted xOy and illustrated in Fig.1. Consequently, twelfth of the structure is enough in order to study the machine\u2019s behaviour (Fig.2). A. 2D flux path 2D flux path consists of the path of the part of the flux created by permanent magnets that participate to the energy conversion (Fig. 3). Its modelling is shown in Fig. 4. Reluctances appearing in the model are indicated in the nomenclature. Fig. 4. Magnetic equivalent circuit for 2D\u2019s path Where : a\u03b5 is the magnetomotive force produced by permanent magnets : 0\u00b5 \u03b5 aerB a \u00d7 = ea is the permanent magnet thickness Br is the residual flux density The total excitation flux is given by : ( ) ( )( ) '22' ' 2 1212341212 1212 RRRRRRR RR ++++ + \u00d7= eqa a v p \u03b5 \u03d5 (1) The reluctances appearing in the previous expression are calculated from the dimensions and the magnetic permeability of each part of the machine [7] and [8] : ( )\u222b \u00d7 = 2 1 l l lS dl \u00b5 R (2) Where : \u00b5 is the magnetic permeability of the concerned part of the circuit l is the length of the flux\u2019s path S is the cross-sectional area of the path The calculation of the reluctance of a half rotor pole is given as an exemple"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002323_978-3-642-25486-4_25-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002323_978-3-642-25486-4_25-Figure2-1.png",
+        "caption": "Fig. 2. Vector and angle definitions within the local reference frame of kinematic chain i (left side). Definition of angle in y-z-plane of local reference frame i.",
+        "texts": [],
+        "surrounding_texts": [
+            "The calculation of the inverse kinematics starts with the vector loop equation for each kinematic chain: .= + + \u22121i 2i i ip l l a b (1) We assume that all joint coordinates of the base frame as well as all joint coordinates of the end-effector lie on circles with radii a and b but have the same angular allocations. In this case, transformation matrices from the global reference frame into the leg-sided reference frames can be introduced [5]: cos sin 0 sin cos 0 . 0 0 1 \u03b1 \u03b1 \u03b1 \u03b1    = \u2212      i i i i i 0T (2) At this point a new auxiliary vector can be calculated: 0 . 0 \u2212 +   = \u2212 + + = + \u22c5      a b i i i i i i i i 0r a p b T p (3) During the inverse kinematic calculation the end-effector position vector p is known. The vector can also be expressed with: ( ) ( ) ( ) ( ) ( ) ( ) ( ) cos sin cos 0 cos . sin sin sin \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5    \u22c5 \u22c5 \u22c5 +    = + = + \u22c5       \u22c5 \u22c5 \u22c5 +    1 1i 2 3i 1i 2i 2 3i 1 1i 2 3i 1i 2i l l l l l i i i i 1i 2ir l l (4) While evaluating the second component equations (3) and (4) lead to the equation for angle : 1 sin cos cos . \u03b1 \u03b1\u03d5 \u2212  \u2212 \u22c5 + \u22c5 =     i i 3i 2 x y l (5) The second auxiliary variable can now be computed with the help of the magnitude of vector : ( ) 2 2 2 2 2 1cos . 2 sin \u03d5 \u03d5 \u2212  + + \u2212 \u2212  =  \u22c5 \u22c5 \u22c5  i i i i,x i,y i,z 1 2 2i 1 2 3i r r r l l l l (6) The actuation angles , which are the final results of the inverse kinematic calculations are also resulting from the evaluation of equations (3) and (4): ( ) sin cos sin sin tan . sin sin sin cos \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u2212 \u22c5 \u2212 \u22c5 \u22c5 \u22c5 + \u22c5 \u22c5 \u22c5 = \u2212 \u22c5 + \u22c5 \u22c5 \u22c5 + \u22c5 \u22c5 \u22c5 i i i 1 i,z 2 3i 2i i,z 2 3i 2i i,x 1i i i i 1 i,x 2 3i 2i i,z 2 3i 2i i,x l r l r l r l r l r l r (7)"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003080_jmer.9000033-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003080_jmer.9000033-Figure4-1.png",
+        "caption": "Figure 4. Force diagram of piston.",
+        "texts": [
+            "G: QQQacx )2cos(4)sin()cos(( 2 21 2 1 \u03c9\u03b8\u03b1\u03b8\u03c9\u03b8 \u2212\u2212\u2212= QQnQ ))2cos())(cos(()2sin(2 542 \u03b8\u03b8\u03b1\u03b1\u03b8 +\u2212\u2212\u2212\u2212 nQQn r sin()cos( 2 1 ))2cos()(sin( 23 54 2 \u03b8\u03c9\u03b8\u03b8\u03c9 \u2212\u2212 COSQQneQ ))2()((cos)sin())2sin( 2 54 222 1 \u03b8\u03b8\u03c9\u03b8\u03b8 \u2212+ iQQe r )))2cos(( 23 \u03b8+ (17) 88 J. Mech. Eng. Res. And for horizontal acceleration of C.G: nQQnacy )(sin())2cos()(cos((( 2 54 \u03b8\u03c9\u03b8\u03b8\u03b1 +\u2212\u2212\u2212= nQQ r ))2sin()cos( 2 1 ))2cos()( 23 54 \u03b8\u03b8\u03c9\u03b8 \u2212\u2212 nQQe )((cos))2cos(( 222 23 \u03b8\u03c9\u03b8 \u2212+ jeQQQ r ))sin())2cos()( 1 2 54 \u03b8\u03b8\u2212 (18) Kinetic analysis of slider-crank mechanism Kinetic calculation must start from the piston because slider-crank mechanism started from that. The force diagram of piston was shown in Figure 4 and Equations 19 and 20: \u2211 = PPx amF . (19) ppgx amFR .\u2212= (20) The force diagram of connecting rod was shown in Figure 5 and Equations 21 and 22:     =\u2212 =\u2211 cxcxx cxcx amRN amF . (21) cxcppxx amamRN .. ++= (22) Engine torque can be obtained from Figure 6 as follow: )cos(..)sin(.. \u03b8\u03b8 rNrNT yx += (23) For Ny: \u2211 = \u03b7.AA IM (24) )2cos(. .)sin(.. 23 \u03b8 \u03b7\u03b8 QQ IrN N Ax y + \u2212 = (25) Where IA is Inertia of connection rod. But for all journals: 4321 TTTTTC +++= (26) That Tc indicates the crankshaft torque, not engine output torque"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003569_iros.2011.6048708-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003569_iros.2011.6048708-Figure4-1.png",
+        "caption": "Fig. 4. Illustration of the snap-in example. The clip is simulated at 1000 Hz whereas the cylinder is simulated at 25 Hz.",
+        "texts": [
+            " First, the validation of action-reaction principle has been performed using two beam models having the identical physical properties, each being calculated at different frequency. The low-rate curve-like object was fixed in the space, whereas the high-rate one was attached to the lowrate body on one end and to the haptic device on the other. Although each model was calculated at different frequency, the deformations of both objects was identical for arbitrary position of the haptic device as shown in Fig. 3. Second, we implemented a snap-in scene, depicted in Fig. 4. In this example, the high-rate object was represented by a stiff but deformable clamp attached to the haptic device via coupling spring. The clip was modeled with 20 corotational beam elements and nodes with positional and rotational degrees of freedom were used in the model. As the obstacle, a deformable cylinder composed of 103 tetrahedral elements was simulated. During the snap operation, up to 20 constraints were created. Various rigidities of both cylinder and clamp were tested with haptic feedback"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000981_s11036-008-0151-4-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000981_s11036-008-0151-4-Figure1-1.png",
+        "caption": "Fig. 1 States of the pursuit and target vehicles. The vector r = [x, y]T is the vehicle position vector",
+        "texts": [
+            " Furthermore, for the sake of simplicity, and because most earth bound vehicles have dynamics that naturally decouple into vertical and horizontal planes, vehicle dynamics will be restricted to a plane perpendicular to gravity. Systems of this sort can be modeled with \u201cunicycle\u201d dynamics, in which the vehicle trajectory is manipulated by heading and speed controls, while the nonholonomic constraint dictates that the vehicle cannot have any component of its velocity in a direction perpendicular to its heading. A model of these vehicle dynamics is shown in Fig. 1, in which r = [x, y]T is the position vector of the vehicle, v is its velocity vector, and \u03b8 is its heading.The unicycle dynamics are typically written in Cartesian coordinates as x\u0307i = vi cos ( \u03b8 i) , i = 1 . . . N y\u0307i = vi sin ( \u03b8 i) \u03b8\u0307 i = ui steer v\u0307i = ui accel, (1) where (see Fig. 1) xi and y i are the Cartesian coordinates of the ith vehicle, \u03b8 i is its heading, vi is its speed, and ui steer and ui accel are its steering and speed controls, respectively. The speed control is included in the model for generality, but in this work the speed control is not used and hence set to zero. These dynamics govern the movement of the pursuit vehicles as they move in the plane. Alternatively, the unicycle model can be written with second order dynamics. This choice is advantageous for the target model as it allows the UKF to estimate the turn rate of the target vehicle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000732_apex.2007.357698-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000732_apex.2007.357698-Figure1-1.png",
+        "caption": "Fig. 1. A typical bearing geometry.",
+        "texts": [
+            " Basically, the middle point velocity, wx, is the mean of inner and outer race linear velocities as given by W = Vx =_ Win rin + Wout rout x 2r , fault current harmonic frequencies are given in (6) employing the previously found mechanical characteristic frequencies. fcf = It \u00b1 mfv (6) where fi is the fundamental stator (carrier) frequency, f, is the characteristic mechanical vibration (modulation) frequency due to bearing fault and m is an integer. (1) where r, is the radius of cage, wi, and w,,, are the angular velocity of inner and outer raceways, rin and r0,, are the radius of inner and outer raceways as shown in Fig. 1, respectively. The respective rotational frequencies are f, ftou and fn, therefore, _ fin rin + fout rout 2r, (2) The outer raceway defect frequency ford is associated with the rate at which the balls pass a defective point on the outer race. Obviously, the frequency increases linearly with the number of balls, therefore the outer race defect frequency is calculated by multiplying the number of balls with the difference of the reference point and the outer race frequencies given by ford = nlfx fout finrin +froutrout -n 2r out(3) =n BDcos/8 2 (fin -fout)(- PD ) where n is the number of balls, /3 is the contact angle, BD and PD are the ball and pitch diameters, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003057_speedam.2012.6264387-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003057_speedam.2012.6264387-Figure1-1.png",
+        "caption": "Fig. 1. Basic motor model (Type A)",
+        "texts": [
+            " The torque ripple is generated by the pulsation of the cogging torque, the variation of the inductance with the rotor position, and the interaction between the magnetmotive force of stator and the magnetic force of the permanent magnet. The rate of the torque ripple rippleT is given as follows: [%]100minmax \u00d7\u2212= ave ripple T TTT (3) where maxT is the maximum torque, minT is the minimum torque, and aveT is the average torque. The cogging torque and the rate of the torque cause the vibration and the acoustic noise. The model structure of the basic IPMSM used in this study is shown in Fig. 1. In this model, the rare earth magnets are used in the rotor, and the stator winding is the distributed one. Table I shows the specifications of the basic motor that is called the basic motor model or Type A in this paper. The rare earth magnets used in Type A are Nd-Fe-B. Torque Characteristics of IPMSM with Spoke and Axial Type Magnets S. Ohira*, N. Hasegawa*, I. Miki*, D. Matsuhashi**, and T. Okitsu** *School of Science and Technology, MEIJI UNIVERSITY, 1-1-1 Higashimita, Tama-ku, Kawasaki, 214-8571 (Japan) **System Technology Research Laboratories, Research & Development Group, Meidensha Corporation, Tokyo 141-6029 (Japan) ( ) qdqdnqan iiLLpipT \u2212+= \u03c8 978-1-4673-1301-8/12/$31"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001110_095308-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001110_095308-Figure1-1.png",
+        "caption": "Figure 1. Schematic of laser sintering process.",
+        "texts": [
+            "6 \u00b5m) laser with a maximum output power of 2000 W, an automatic powder delivery system and a computer for process control. When a specimen was prepared, a steel substrate was placed on the building platform and levelled. Afterwards, a thin layer of loose powder (0.20 mm in thickness) was deposited and levelled on the substrate by the roller. Subsequently, a laser beam was used to scan the powder bed surface to form a layerwise profile according to CAD data of the specimen. The used scanning pattern is shown schematically in figure 1: a simple linear raster scan with a short scan vector. The process was repeated until the specimen was completed. The following suitable processing parameters were used: spot size of 0.30 mm, laser power of 700 W, scan speed of 0.06 m s\u22121 and scan line spacing of 0.15 mm. Rectangular test specimens with dimensions of 50 \u00d7 10 \u00d7 6 mm3 were produced. The densities of sintered specimens were measured using Archimedes\u2019 method. Samples for metallographic examinations were cut from as-sintered specimens and sequentially plane ground with SiC sandpaper to a 1000 grit finish"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure4-1.png",
+        "caption": "FIG. 4. Four-pad hydrostatic journal configuration.",
+        "texts": [
+            " Loading the spindle vertically upwards in the x\u00b1 direction produces a deflection in the y direction and corresponds to a stiffness Xyxi = (\u2014 ) 470,000 lbf. The actual temperature inside the bearing was not assessed but the oil temperature in proximity to the bearing reached a maximum of 80\u00b0F at 23 rev/min and 108\u00b0F at 385 rev/min. 2. HYDROSTATIC OIL BEARING CONTROL WHEEL SPINDLE UNIT The working principle of hydrostatic bearings is now well known. <2> The support medium, in this case oil, is passed under pressure through a restrictor before entering bearing pockets (Fig. 4). Lands at the boundaries of the pockets cause further restriction and the medium is then exhausted at low pressure. Variations of the bearing gap cause changes in the resistance to flow and hence a compensating change of pressure occurs in the bearing pocket. The degree of pressure compensation depends greatly on the type of flow restriction. However it may be shown that for capillary restrictors or other devices where restriction is directly proportional to viscosity, temperature does not affect the bearing stiffness, with the only proviso that the restrictor must be maintained at the same temperature as the bearing lands. This characteristic does not apply for orifice compensated bearings. When designing journal bearings it is important to calculate the following values as described in the Appendix. W. B. ROWE (i) Oil flow is the unloaded condition go For the configuration shown in Fig. 4, oil exhausts axially, so that for the unloaded condition the circumferential flow resistance is not a parameter. The flow may be calculated from: go = s go (in3/sec) Ps is the supply pressure (lbf/in2), ho is the bearing gap (in.), \u03bc is the dynamic viscosity (reyns). The non-dimensionalized flow is : go = \u03b7\u0392\u03b2 (dimensionless) where n is the number of bearing pockets, B is the shape factor = , 6nti \u03b2 is the ratio of pocket pressure to supply pressure and for most applications has an optimum values close to 0-5",
+            ") ho where A'e = D(L-h) (in2) is the effective bearing area related to: \u03bb\u03bf = non-dimensionalized stiffness. The non-dimensionalized stiffness takes account of considerations of bearing geometry, number of bearing pockets, and circumferential flow. A circumferential flow factor y is introduced in the calculation to provide for the deleterious 'effect [on stiffness of flow from heavily loaded pockets into lightly loaded pockets, and is the ratio of axial flow resistance to circumferential flow resistance in the unloaded condition. For the configuration of Fig. 4 y = \u2014-\u2014 (dimensionless). TTD \u00ce2 The non-dimensionalized stiffness is : 3-82/3(1-0) A \u00b0 - i + r ( i - / 0 458 W. B. ROWE which may be optimized for stiffness when Poptimum \u2014 (y + i)\u00e8 1 + (y + 1)* However, if \u00df = 0-5, \u03b3 = 0-5, then \u03bb0 = 0-76. (iv) Lead-bearing capacity WmSiX ^ m a x \u2014 ^ m a x P s A e where WmBlX = 0-707 (Pi,max \u2014 ^3,min) is the non-dimensionalized load-bearing capacity. l + 0 - 0 2 5 y ( l - j S ) \u2022*i\u00bbmax \u2014 / . o\\ 1 + l - P) (0-004 + 0-025y) P Pz, l + ( 4 - 9 7 y l - | 8 ) min \u2014 Q n\\ 1 + ~ \u2014 -- (6-9 + 4-97y) The bearing will usually have been designed for the smallest possible bearing gap consistent with economic machining tolerances in order to maximize stiffness",
+            " If it is assumed that oil is supplied from a source at ambient temperature, the heat dissipated in the bearing is the sum of the pumping power and the friction power resulting from relative sliding of the two bearing members : Ideal pumping power Np = PSQ (lbf in. sec\"1) Friction power NF = \u03bc\u03a521 ~ + \u2014 I (lbf in. sec-1) Grinding Machine Spindles 473 An over-estimate of the temperature rise of the oil flowing through the bearings is \u0394\u0393=(^\u00b1^\u03bf (\u03b2\u03c1) (A2) IM) {Jo for a range of oils such as Shell Tellus 27. (a) Journal Bearing with Four Pads\u00ae) Fig. 4. Capillary compensated Equating the flow through the restrictor to the flow through the bearing gi = \"dZ\u00b0 - Pl) \u00b0 - ' \u00b7 ft + \u0302 ^ (ft - Pa) (A3) 1 2 \u00f6 / x / c /x /x If the land width is much smaller than the recess width the shape factor is 6nt\u00b1 Allowance for circumferential flow is introduced by the factor y, the rate of axial flow resistance to circumferential flow resistance in the unloaded condition. Letting in eqn. (A3) Hence and \u03b7\u03b9 = \u03b72 = h0; Pi=P2=Po = \u00dfPs 7\u0393\u03b9 d4 _ \u00df Bhl 128ft lc 1 - \u03b2 \u03bc \u00dfo - nB\u00df \u00c4 f t + ri1(ft_ft,__L(P"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.7-1.png",
+        "caption": "Fig. 1.7. The influence of load transfer on the resulting axle characteristic.",
+        "texts": [
+            "8) and (1.13) non-linear curves may be adopted, possibly obtained from measurements. For the case of roll camber, the situation becomes more complex. 14 TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY At a given axle side force the roll angle and the associated camber angle can be found. The cornering characteristic of the pair of tyres at that camber angle is needed to find the slip angle belonging to the side force considered. Load transfer is another example that is less easy to handle. In Fig. 1.7 a three dimensional graph is presented for the variation of the side force of an individual tyre as a function of the slip angle and of the vertical load. The former at a given load and the latter at a given slip angle. The diagram illustrates that at load transfer the outer tyre exhibiting a larger load will generate a larger side force than the inner tyre. Because of the non-linear degressive Fy vs F z curve, however, the average side force will be smaller than the original value it had in the absence of load transfer"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003080_jmer.9000033-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003080_jmer.9000033-Figure5-1.png",
+        "caption": "Figure 5. Force diagram of connecting rod.",
+        "texts": [
+            " And for horizontal acceleration of C.G: nQQnacy )(sin())2cos()(cos((( 2 54 \u03b8\u03c9\u03b8\u03b8\u03b1 +\u2212\u2212\u2212= nQQ r ))2sin()cos( 2 1 ))2cos()( 23 54 \u03b8\u03b8\u03c9\u03b8 \u2212\u2212 nQQe )((cos))2cos(( 222 23 \u03b8\u03c9\u03b8 \u2212+ jeQQQ r ))sin())2cos()( 1 2 54 \u03b8\u03b8\u2212 (18) Kinetic analysis of slider-crank mechanism Kinetic calculation must start from the piston because slider-crank mechanism started from that. The force diagram of piston was shown in Figure 4 and Equations 19 and 20: \u2211 = PPx amF . (19) ppgx amFR .\u2212= (20) The force diagram of connecting rod was shown in Figure 5 and Equations 21 and 22:     =\u2212 =\u2211 cxcxx cxcx amRN amF . (21) cxcppxx amamRN .. ++= (22) Engine torque can be obtained from Figure 6 as follow: )cos(..)sin(.. \u03b8\u03b8 rNrNT yx += (23) For Ny: \u2211 = \u03b7.AA IM (24) )2cos(. .)sin(.. 23 \u03b8 \u03b7\u03b8 QQ IrN N Ax y + \u2212 = (25) Where IA is Inertia of connection rod. But for all journals: 4321 TTTTTC +++= (26) That Tc indicates the crankshaft torque, not engine output torque. Note that the friction force is negligible in comparison with gas force, so it has ignored in calculations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure2.19-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure2.19-1.png",
+        "caption": "Figure 2.19 Normalized response vectors in the complex plane",
+        "texts": [
+            " 42 Closing the Loop The \u2018amplitude ratio\u2019 of the response is the ratio of the lengths (moduli) of the vectors and the phase shift is the phase angle difference between the two vectors, i.e. amplitude ratio (AR)= \u2223\u2223\u2223\u2223xoxi \u2223\u2223\u2223\u2223 phase angle = \u2220xi\u2212\u2220xo To summarize, the frequency response can be calculated by following the simple rules below: 1. Substitute D = j in the transfer function. 2. Plug in values of covering the frequency range of interest. 3. Gather up the real and imaginary terms at each frequency. If we assume that the forcing function vector always lies along the positive real axis in the complex plane as indicated by the diagram of Figure 2.19 we can say: Amplitude ratio= \u221a R2+ I2m Phase shift = tan\u22121 ( Im R ) Here R and Im are the real and imaginary components of the output vector respectively. Calculating Frequency Response 43 Having developed the basic rules for calculating the amplitude ratio (or gain) and phase angle of transfer functions, let us now calculate the frequency response of the generic first-order lag defined as follows: 1 1+ j T Based on the rules developed above we can say: We can now plot the results for gain and phase angle against frequency as shown in Figure 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.28-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.28-1.png",
+        "caption": "Fig. 1.28. The MTS Flat-Trac Roadway Simulator TM, Milliken (1995).",
+        "texts": [
+            " Another more practical solution would be to bring the vehicle in the desired attitude (fl--8 ~ by briefly inducing large brake or drive slip at the rear that lowers the cornering force and lets the car swing to the desired slip angle while at the same time the steering wheel is turned backwards to even negative values. The MMM diagram, which is actually a Gough plot (for a single tyre, cf. Figs.3.5 and 3.29) established for the whole car at different steer angles, may be assessed experimentally either through outdoor or indoor experiments. On the proving ground a vehicle may be attached at the side of a heavy truck or railway vehicle and set at different slip angles while the force and moment are being measured (tethered testing), cf. Milliken (1995). Figure 1.28 depicts the remarkable laboratory MMM test machine. This MTS Flat-Trac Roadway Simulator TM uses four flat belts which can be steered and driven independently. The car is constrained in its centre of gravity but is free to roll and pitch. 1.3.6. The Car-Trai ler Combination In this section we will discuss the role of the tyre in connection with the dynamic behaviour of a car that tows a trailer. More specifically, we will study the possible unstable motions that may show up with such a combination"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001086_s11044-007-9053-7-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001086_s11044-007-9053-7-Figure1-1.png",
+        "caption": "Fig. 1 The angles yi of the Andrews\u2019 system",
+        "texts": [
+            "eywords Constraint singularities \u00b7 Andrews squeezing mechanism \u00b7 Ideal decomposition \u00b7 Gr\u00f6bner bases \u00b7 Descriptor form \u00b7 Angular coordinates The \u201cAndrews\u2019 squeezing system\u201d was first described by Giles in [7] and further studied in [12]. It is a planar multibody system whose topology consists of closed kinematic loops (see Fig. 1). The Andrews\u2019 system was promoted in [17] as a benchmark problem to compare different multibody solvers. Nowadays it is a well-known benchmark problem [10, 14] for numerical integration of differential-algebraic equations as well. The equations are of the T. Arponen ( ) Institute of Mathematics, Helsinki University of Technology, PL 1100, 02015 TKK, Helsinki, Finland e-mail: teijo.arponen@hut.fi S. Piipponen \u00b7 J. Tuomela Department of Physics and Mathematics, University of Joensuu, PL 111, 80101 Joensuu, Finland S"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002928_elps.201100355-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002928_elps.201100355-Figure1-1.png",
+        "caption": "Figure 1. Fabrication steps and photograph of tube-like microchannels patterned on a flat surface. (A) The master pattern is obtained by a two-level photolithography with SU8 3050 photoresist. The pattern is then replicated into a PDMS layer by casting. Finally, the PDMS mold is used to structure a tube-like channel by UV-assisted molding and irreversible bonding onto the surface of a glass slide. (B) An array of tube-like channels filled of different colors.",
+        "texts": [
+            "4, Type VII, from Aspergillus Niger), n-hexane, D-(1)-glucose, bovine serum albumin (BSA), glutaraldehyde (GA, 25% aqueous solution), potassium iodide (KI), iodine (I2), sodium chloride (NaCl), sodium phosphate monobasic (NaH2PO4), sodium phosphate dibasic (Na2HPO4), sulfuric acid (H2SO4, 98%), hydrogen peroxide (H2O2), and trimethylchlorosilane were purchased from Sigma (France). 2-Mercaptoethanol was purchased from Acros Organics (France). p-Benzoquinone (PBQ) was purchased from Fluka (France) and recrystallized from n-hexane. All solutions were prepared with deionized water (18.2 MO/cm). An aliquot of 0.1 M PBS consisted of 0.05 M NaCl and 0.05 M NaH2PO4/Na2HPO4, and the pH was adjusted to 7.3 and 5.6, respectively. In total, 0.1 M glucose solution was prepared in 0.1 M PBS (pH 7.3). Figure 1A shows schematic diagram of the device fabrication steps. First, a 50-mm thick SU8 3050 resist layer was spin-coated on a clean silicon wafer. After UV exposure and development, a second-level photolithography was applied with the same SU8 3050 resist but different mask features. Afterward, trimethylchlorosilane was evaporated on the SU8 sample as release agent. Prepolymer solution of PDMS was prepared at a 10:1 ratio of base polymer to cross-linker with a Cyclone UM-103S mixer (UNIX, Japan). Then, it was poured on the two-level SU8 structures",
+            " After curing at 801C for 2 h, the cross-linked PDMS layers were peeled off and the structured PDMS layer was placed on the flat PDMS, forming an enclosed mold with inlets. After degassing of the PDMS assembly, NOA 81 was filled in by gas reabsorption assisted self-loading, followed by a UV exposure (12 s, 30 J/cm2). Here, the exposure dose was optimized to be sufficient for the structure precuring but this precured structure can still be not exposed for further processing. Afterward, the PDMS mold was peeled off and resulted tube-like structure was bonded onto a glass substrate by additional UV exposing (70 s, 30 J/cm2). Figure 1B shows a fabricated microtube array filled of liquid of different colors. This array has been designed for eye view with 50 500 mm2 inner cross-section and 100 2000 mm2 outer cross-section. Much smaller tube-like microchannels could also be obtained with the same fabricated process. Finally, a thick PDMS block containing microvalves or pumps could be added by placing it on the top of microtube arrays with access holes. To avoid leakage, this PDMS block could be irreversibly bonded on the microchannel device by curing at 801C overnight",
+            " Due to its high elastic modulus, NOA 81-defined channels could be made with large channel widths without structure collapse. Comparing to the similar structure fabricated by PDMS device, the device made of NOA 81 is impermeable to gas and water vapor, so that it prevents evaporation of liquids to assure long-time measurements with very small sample volumes. The surface of cured NOA 81 is natively hydrophilic so that dye-doped liquids could be easily injected into the microchannels by simply depositing a droplet in the entrance of the channels (Fig. 1B). Besides, the devices made of NOA 81 conserve the most of advantages of the similar devices made of PDMS such as large open-access areas outside the channels for cell-based assays. Finally, thick PDMS blocks could be easily added for the integration of more conventional microfluidic functionalities for glucose analysis (Fig. 2). After assembling, the fluidic and electric connections could then be done with outward apparatus. The flip-working electrode configuration takes another advantage of tube-like microchannels for electrochemical detection with disposable enzyme-working electrodes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001868_1.3591482-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001868_1.3591482-Figure1-1.png",
+        "caption": "Fig. 1 A section of the general spatial kinematic chain showing the Icth bar-slideball member ball-jointed to adjacent members. The vector Aa-, representing the fcth member in the 1st position, stretch-rotates to A i n the/th position.",
+        "texts": [
+            " With this approach and by taking derivatives, first, second, and higher-order loop equations can be developed which form the basis for a general method of spatial kinematic synthesis, applicable to path, function and motion generation (body guidance) with first, second, and higher-order as well as for combined \"point-order\" ap proximo lions. Introduction IN a recent paper [l]1 a general spatial kinematic chain was introduced consistingof one or more interconnected loops of successively ball-jointed bar-slideball members (Fig. 1). In a starting or reference position, each such member can be represented by a vector Au . , where the subscript 1 designates the first position and It the fcth member. In general, in the jih position, the vector will have stretch-rotated to become A,/.-. In reference [1] the stretch-rotation is accomplished by means of a quaternion operator [6, 7, 21]. Matrix methods [12, 22, 23] inspired the present paper, leading to a stretch-rotation tensor which may lie introduced as follows. 2 The Stretch-Rotation Tensor Let a vector R be operated on by an orthonormal rotational transformation tensor T = [/l7] [2], such that = Sjt, j,k = 1 , 2 , 3 (1) where, according to standard indicial notation, summation over repeated indexes is understood, and where Sjk",
+            "] for the stretch-rotation from R to pR', it is easy to see that the inverse stretchrotation from pR' to R is obtained by means of the operator [pT\\ = p - ' T \" 1 = p-'T', (25) where we made use of the fact that the inverse and the transpose of an orthogonal matrix are equal. Thus, for the example of Fig. 2, [p2']- ' = p- 'T< = 'A l o 0 0 0 - 1 0 0 - ' A which, operating on pR' = (2, 0, 2), yields R = (1, 1, 0). 5 Spatial Loop Equations The closure of each spatial loop in the j'th position of the chain (Fig. 1) is expressed by the vector equation: E = o. A = 1 (26) where n is the number of bar-slideball members in the loop. In terms of the stretch-rotation tensor we can write this equation as follows: E p>A-iu^AU- = o, i = l (27) where we use the indexes u and v for the tensor components in order to distinguish them from the position and member subscripts j and k. Equation (27) represents three scalar equations as follows: E Pit E ( O y t ( A . * ) . = 0, u = 1, 2, 3 (28 ) k = 1 o = l where (Au-)\u00bb is the wth coordinate2 of the vector An-"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001321_j.mechmachtheory.2007.08.001-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001321_j.mechmachtheory.2007.08.001-Figure1-1.png",
+        "caption": "Fig. 1. Parallel manipulator: PCM (from [1]).",
+        "texts": [
+            " Finally, this method has other advantages, and it can be even used to determine the existence of inactive joints and equivalent serial chain of complicated parallel chain. The paper is organized as follows: Section 2 discusses the degree-of-freedom calculation of simple parallel mechanisms. Equivalent serial chains of complicated parallel chains are studied in Section 3. Section 4 is about an example of complicated parallel mechanism: Delta robot. Section 5 discusses some interesting applications of this method. Finally, the conclusions are given in Section 6. For the parallel mechanism (PCM) in Fig. 1, axes of same chain are parallel and those of different chains are vertical. All chains are serial mechanisms, so it belongs to simple parallel mechanisms. Obviously, there is J A J B 0 J A 0 J C _Q \u00bc 0 \u00f01\u00de where, JA, JB and JC are Jacobian matrixes of chain A, B and C, respectively. Q is the N \u00b7 1 vector of all joint values [21]. If the coordinates of centers of translational joint in chain A are (2, 0.25, 0.7) and those of three revolute joints are (2, 0.25, 0.5), (1.5, 0.5, 0) and (1,0,0) in turn, then Jacobian JA is: J A \u00bc 0 1 1 1 0 0 0 0 0 0 0 0 1 0 0 0 0 0:5 0 0 0 0:25 0:5 0 2 666666664 3 777777775 \u00f02\u00de Similarly, if the coordinates of centers of joints in chain B are ( 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure6-1.png",
+        "caption": "Fig. 6 Normal and modified gear profile models: (a) standard tooth, (b) modified tooth",
+        "texts": [
+            " Therefore, load distribution along the meshing line is not constant, but changes step by step, and force distribution along the profile depends on contact ratio. As the contact ratio approaches 1, the ratio of the length of a single meshing area to that of a double meshing area increases, as illustrated in Fig. 5(b). 3.2 Procedure of width modification With this type of modification, tooth width, b, has been widened in order to maintain a constant ratio of gear load to gear width (F/b), which tends to vary continuously along the profile of a standard gear. In Fig. 6, a normal and a modified gear model are shown along with their respective load distributions in Figs. 7 and 8. Face width, b, is not standardized, but generally, 9 m < b < 14 m. The wider the face width, the more difficult it is to manufacture and mount the gears so that contact is uniform across the full face width [20]. However, where the volume is limited (for instance, in gearboxes), then recommended values lie in the range 6 m < b < 12 m. If the given values are considered, then width limits in gear width modification will remain within acceptable standards",
+            " That is, flash temperature values between points DC and CB will fall by 40 to 50%. Also, the shaded area in the flash temperature diagram (Fig. 12) will disappear. Dissipation of heat from gears, apart from conditions like surrounding medium temperature and thermal conductivity, is also closely dependent on gear tooth surface area. In addition, with this modification, the amount of heat dissipated will increase and the modification will help to create equilibrium between heat generation and heat dissipation. In measurements made on gears given in Fig. 6, an increase of 20% in the total surface areas of modified gears was found. Therefore, as seen in Fig. 13, this area increase has positive effects in dissipating heat away from the gear tooth. With the width modification made along the meshing line, gear damages are expected to decrease, especially those caused by Hertz surface pressure. Surface contact fatigue of gear teeth is one of the most common causes of gear operational failure due to excessive local Hertzian contact fatigue stresses. Generally, there are two types of surface contact fatigue, namely, pitting and spalling"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002826_bf01690470-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002826_bf01690470-Figure1-1.png",
+        "caption": "Fig. 1. The point force near the grain boundary.",
+        "texts": [
+            " For this kind of solution it was sufficient to know the components of displacement f rom the point force (in the Kelvin sense) in an infinite isotropic medium. The same method can be used even when solving the interaction of the infinitesimal loop with the boundary of two isotropic media. According to H e a d [16] a welded or slipping boundary for the model of the grain boundary is used. The expressions for the displacement field f rom the point force in the respective eases o f welded and slipping boundaries were calculated by R o n g v e d [8], D u n d u r s and H e t d n y i [9]. We shall make use of the following notat ion: (see Fig. 1.) u x, ur, u z (resp. u i) k'l ' ~'2 G1, G 2 x ' , y ' , z ' (resp. x~) x, y, z (resp. xi) Fx, Fy, F z (resp. F i) Components of the displacement field from the point force. Poisson constants for ha/f-spaces I) and 2). Shear moduli. Coordinates of point force and of infinitesimal loop respectively, see Fig. 1. Coordinates of some point in the space. Components of point force. The interface between two solids is the plane z = 0. We shall further introduce: (1) Ul = 3 - 4v 1 , '~2 = 3 - 4 v 2 , / \" - G 2 , M - 1 G 1 4 n G l ( t q + 1) A = 1 - F B - x 2 - K1F S = - - 1 - F 1 + ~ l F ' x 2 + F ' 1 + F ' D = (~, + 1 ) r , (0____o<__1). (~c l + 1) F + t r 2 + 1 (2) R1 = x/l-(x _ x,)2 + (y _ y,)2 + (z - z') 2] R2 = x/[(x - x') 2 + (Y - y,)2 + (z + z')2] . Czech. J. Phys. B 20 (1970) 703 I. Vayera Then we can write the expressions of Rongved in a rather simpler form while using the notation from the work by D u n d u r s and H e t 6 n y i [9]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001456_cca.2007.4389321-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001456_cca.2007.4389321-Figure1-1.png",
+        "caption": "Fig. 1. Geometry and axes definition.",
+        "texts": [
+            " In [8] a discrete adaptive control strategy for coordinated control of AUVs is presented. In [9] a variety of control methodlogies, including sliding mode control, adaptive control and output feedback control of AUVs are given. This paper adresses the problem of tracking the position and attitude of an AUV in the horizontal plane, using two rudders in front and rare side of the vehicle. An adaptive control law is presented to effectively compensate the hydrodynamic effects, in the presence of external disturbances. A schematic diagram of the system under consideration is shown in Fig. 1. Organization of the paper: Introductory materials and a review of the current research trends in literature are treated in section I. Section II treats the dynamic modeling of the system under consideration. Parameters that include uncertainties are introduced in this section. The feedback linearization control law is presented in section III-A. In section III-B, the adaptation law and control law are derived and parameter uncertainties are taken into account in the control law. Section IV introduces the type of the ocean current disturbances that will be used in simulations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002546_s1052618810050018-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002546_s1052618810050018-Figure1-1.png",
+        "caption": "Fig. 1. Mechanism with four degrees of freedom and kinematic decoupling: (a) scheme of the mechanism; (b) system of kinematic screws in the usual configuration; (c) system of kinematic screws in the special provision.",
+        "texts": [
+            " One of them is that all the connecting kinematic chains impose the same relations corresponding to a four component closed group of screws [3, 4]. Another method is connected with the fact that one or more kinematic chains impose a connection corresponding to that group, and other chains do not impose the connections or correspond to a smaller number of connections. Consider the scheme of mechanisms of parallel structure with four degrees of freedom obtained by these methods\u2014three translational movements and rotation around the axis. Let two kinematic chains be imposed on one connection and a third chain make two connections (Fig. 1). The first and the second kinematic chain consist of one drive translational pair (linear motor) located on the base, three intermediate rotational pairs arranged with axes parallel to the axis of linear actuator, and the final rotational pair (the axis of the end rotational pairs of the two chains are the same). The third kinematic chain contains one rotational drive pair (angular motor) mounted on the base, one drive translational pair (the axis of the two pairs overlap), and two translational pairs made in the form of hinged parallelograms (Fig. 1a). Single screws characterizing the status of the axes of these kinematic pairs have the coordinates E11(0, 0, 0, 1, 0, 0), E12(1, 0, 0, 0, , ), E13(1, 0, 0, 0, , ), E14(0, 0, 1, ,e12y 0 e12z 0 e13y 0 e13z 0 e14x 0 DOI: 10.3103/S1052618810050018 408 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 39 No. 5 2010 GLAZUNOV et al. , 0), E21(0, 0, 0, 0, 1, 0), E22(0, 0, 0, , 0, ), E23(0, 0, 0, , 0, ), E24(0, 0, 1, , , 0), E31(0, 0, 1, 0, 0, 0), E32(0, 0, 0, 0, 0, 1), E33(0, 0, 0, , , 0), E34(0, 0, 0, , , 0). Screws E11, E21, E32, E33, and E34 have an infinitely large parameter. The parameter of the remaining screws is equal to zero. The first and second kinematic chains are imposed on one connection, and the third circuit corresponds to two connections. The connections of the first and second kinematic chains can be repeated, because they do not affect the total number of degrees of freedom, equal to four. The power screws of connections imposed by kinematic chains have the coordinates (Fig. 1b) R2(0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0). All the kinematic screws of motion of the output link can be represented as screws mutual to the power screws: \u21261(0, 0, 0, 1, 0, 0), \u21262(0, 0, 0, 0, 1, 0), \u21263(0, 0, 0, 0, 0, 1), \u21264(0, 0, 1, 0, 0, 0); the screws \u21261, \u21262, and \u21263 have an infinitely large parameter, and the screw \u21264 has a zero parameter. Special provisions relating to the loss of one or more degrees of freedom arise if the kinematic screws corresponding to the unit vectors Ei2, Ei3, Ei4 (i = 1, 2) or to the unit vectors E33, E34 are linearly dependent. This occurs if any three screws Ei2, Ei3, Ei4 (i = 1, 2) fall in one plane or if the two screws E33, E34 are par allel. In particular, if any three screws Ei2, Ei3, Ei4 (i = 1, 2) fall in one vertical plane, there exist three power screws imposed by kinematic chains R1(0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0), and R3(0, 0, 1, 0, 0, 0) (Fig. 1c) and only three kinematic screws of the movement of the output link mutual to these screws \u21261(0, 0, 0, 1, 0, 0), \u21262(0, 0, 0, 0, 1, 0), and \u21263(0, 0, 1, 0, 0, 0). Note that R3 is located on the axis z. If all drives are fixed, then there are six force screws imposed by kinematic chains: R1 (0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0), R3(1, 0, 0, 0, 0, 0), R4(0, 1, 0, 0, 0, 0), R5(0, 0, 1, 0, 0, 0), and R6(0, 0, 0, 0, 0, 1). The screws R3, R4, and R5 have a zero parameter, and the screws R1, R2, and R5 have an infinitely large param eter"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000447_14399776.2008.10781294-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000447_14399776.2008.10781294-Figure3-1.png",
+        "caption": "Fig. 3: Sectional view sketch of typical spool valve with main mechanical elements of the proportional valve with input dead zone",
+        "texts": [
+            " 1: Graphical representation of the dead zone The dead zone analytical expression is given by Eq. 1. zm ( ( ) ) ( ) ( ) 0 ( ) ( ( ) ) ( ) md u t zmd if u t zmd u t if zme u t zmd me u t zme if u t zme \u2212 \u2265\u23a7 \u23aa = < <\u23a8 \u23aa \u2212 \u2264\u23a9 (1) Pneumatic Valves In proportional pneumatic valves of directional control, dead zone is located at the dynamic system as a block diagram shown in Fig. 2. To understand this phenomenon better will be presented a detailed depic- tion of three-lands-five-ways spool valve components and its working. Figure 3 shows a sectional view sketch of a typical spool valve with main mechanical elements that can be used as a proportional valve. The control signal u energizes valve\u2019s solenoids so that a resulting magnetic force is applied in the valve\u2019s spool. An example of industrial spool proportional pneumatic valve is Festo MPYE-5 (Festo, 1996). In closed centre or overlapped valves, the land width is greater than the port width when the spool is at null position, resulting the presence of the dead zone nonlinearity (Merrit, 1967)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000772_robio.2007.4522451-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000772_robio.2007.4522451-Figure1-1.png",
+        "caption": "Fig. 1 This concept wall-climbing robot for scout task The Fig.2 is the robot suction principle , It is mainly composed of suction cup, flexible sealed ring and propeller.",
+        "texts": [
+            " In this paper, we use fluid network theory to set up a dynamic model and found the dynamic response equation of the negative pressure, and then simulate the transition process in order to see the responsive time of the pressure in the suction cup. The effect that the key parameters of the suction cup have on the suction characteristic will be analysed in the following. Finally, in the experiment section, the results for adsorbing on various kinds of wall surfaces are discussed. Aim at the serious situations for anti-terror in the modern society, we presents a small-sized wall climbing robot using this novel adsorption method (CSM), Fig.1 is to express the robot concept equipped with a camera manipulator. Once propeller goes round and round at full speed, the air vents and thrust force produces that pushes the suction cup to the wall. What\u2019s more, air enters into the suction cup through the flexible sealed ring and it makes the cup achieve the negative pressure state. So there is the pressure force for the robot to suck on the wall. By adjusting the gap between the 978-1-4244-1758-2/08/$25.00 \u00a9 2008 IEEE. 1866 Proceedings of the 2007 IEEE International Conference on Robotics and Biomimetics December 15 -18, 2007, Sanya, China sealed ring and the wall surface, the critical suction would be obtained in the robot suction system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001077_detc2009-87092-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001077_detc2009-87092-Figure4-1.png",
+        "caption": "Figure 4. Mesh generation",
+        "texts": [
+            "2 N/mm3 at rated torque. Because the cycloid disk is expected to move due to finite bearing stiffness and contact force distribution will change significantly, spring elements for all 40 contact points are included initially, and spring elements that are supposed to have tensile forces are removed through iterative FE analysis. Triangular meshes of 1 mm are automatically generated for the cycloid disk, and fine sphere and mapped meshes of half that size are built for the contact region, as shown in Figure 4. (Node #: 34337, Element #: 97365). Theory: Hertz contact theory is one of the most classical and fundamental theories concerning contact mechanics of elastic solids [7]. According to Hertz contact theory, th contact stiffness, kH, which is contact force per unit contact depth, can be expressed as a nonlinear function of contact force, material properties, and geometries of two contact cylinders, as shown in Eq. (3). Variables for Eq. (3) are described in Table 2. 2 2 1 1 2 2 1 2 1 4 1 1 4 12 ln ln 2 2 H Wk R R E a E a \u03c0 \u03bd \u03bd = \u23a1 \u23a4\u2212 \u2212\u239b \u239e \u239b \u239e\u2212 + \u2212\u239c \u239f \u239c \u239f\u23a2 \u23a5\u239d \u23a0 \u239d \u23a0\u23a3 \u23a6 (3) Here 1 2*2( ) * PRa E W\u03c0 = Downloaded From: http://proceedings"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure5-1.png",
+        "caption": "Fig. 5. Bending loadcase",
+        "texts": [
+            " The FBS technique can be used to convert the free-free system to a constrained system: \u2022 subsystem A: free-free FRF's of BIW as shown in Fig. 2 \u2022 subsystem B: ground To represent the static test bench condition for torsional stiffness, the rear domes will be grounded (Fig. 3). To represent the static test bench condition for bending stiffness, rear and front domes will be grounded (Fig. 4). Forced Response The bending stiffness of a vehicle body is measured by clamping the body at the four domes and applying a load at the 4 seat bolting positions (F1 to F4) as is represented in Fig. 5. (2) After calculating the coupled system matrix Hc from equation (1) the displacement in the output points d(\u03c9) can be easily calculated using a forced response described in equation (2). This is a general approach which could be used for several load cases. The forced response can be done over a certain frequency range (e.g.: 5-30 Hz) and the static stiffness can be extracted by fitting/extrapolating the curve towards 0 Hz. This fitting will be necessary since at the very low frequencies, the results will be very noisy when working with measured FRF data"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000816_icma.2008.4798779-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000816_icma.2008.4798779-Figure3-1.png",
+        "caption": "Fig. 3 Geometrical model for supporting forward tilting posture",
+        "texts": [
+            " Waist belt is used as a normal belt for pants. When the McKibben artificial muscle is constricted, inner black circle is forced to rotate with respect to back frame. Waist frame directly connected to inner black circle is applied and by using thigh belt through connecting belt, waist frame is fixed in order to avoid rotation. In this mechanism, the front part of thigh is burdened by thigh belt during forward tilting support. Let\u2019s here investigate how much load the front part of thigh has. Geometrical model is shown in Fig.3. O describes the center of rotation at waist. AT [Nm] denotes torque by the upper body as shown in eq.(1), \u03b8sinMgTA \u22c5= (1) where M [kg] and [m] indicate mass of the upper body and distance from the center of rotation to the gravity center of the upper body. BT [Nm] depicts torque opposes AT as described in eq.(2), tB TbT \u03b8cos\u22c5= . (2) If AB TT \u2265 , waist belt stays without revolving. Here let\u2019s calculate T when AB TT = , i.e., t t b MgT TbMg \u03b8 \u03b8 \u03b8\u03b8 cos sin cossin \u22c5=\u2234 \u22c5=\u22c5 (3) The load for horizontal and vertical direction of the front part of thigh ( xT and yT respectively) are as follows; \u03b8\u03b8 \u03b8 \u03b8\u03b8 sinsin cos sinsin Mg cb MgTT t t tx =\u22c5== (4) \u03b8\u03b8 \u03b8 \u03b8\u03b8 sincos cos sincos Mg bb MgTT t t ty =\u22c5== (5) As a result, one leg receives 2xT and 2yT "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001918_j.aquaeng.2009.06.002-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001918_j.aquaeng.2009.06.002-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of the static PECO test system.",
+        "texts": [
+            "92 software (PAR), which also recorded the current and voltage. The photoanode was rolled into a cylinder and placed against the inner wall of a 300-mL glass beaker. A 9-W germicidal UV lamp was positioned within a quartz sleeve (32 mm ID, 35 mm OD, 15 cm long), in the center of the beaker. The distance from the lamp to the photoanode was approximately 2 cm. The cathode and reference electrode were attached to the outer wall of the quartz sleeve and positioned parallel to each other 2 cm apart (Fig. 1). Glass beakers were filled with 250 mL of either 100% seawater made using 40.5 g/L Instant Ocean1 or water containing 1 g/L NaCl (low-salt water). The water was spiked with ammonium chloride to provide an initial concentration of 0.5\u2013 9 mg/L nitrogen as ammonia. The wetted area of each photoanode was 180 cm2. In most experiments, air was bubbled into each beaker to provide aeration and uniform mixing of the water. All experiments were conducted at room temperature (22 8C 2 8C). During studies assessing the effects of pH on ammonia removal in low-salt water (1 g/L NaCl), pH was adjusted using NaOH or HCl and measured with a pH meter"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003146_amr.383-390.6242-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003146_amr.383-390.6242-Figure1-1.png",
+        "caption": "Fig. 1 Photograph of the DMLS components Fig. 2 Experimental Setup",
+        "texts": [
+            " In this study the components were manufactured by optimized value of process parameters obtained using multi-objective optimization techniques. The optimized values are sintering speed (250 mm/s), hatch spacing (0.25mm), hatch type (shifted), post contouring speed (350 mm/s) and infiltration using epoxy resin. Standard components used in an automobile starter motor were used as reference for 3D modeling. Bushes were manufactured as per CAD data. Surface roughness (Ra) was measured for all components before functional testing. Photographs of manufactured component are shown in fig.1. Manufactured components were tested for real time application to study the wear behavior. As per the functional requirements, all components were tested. Given below are the details of components used for testing. An automobile self starter motor was selected for testing. In actual working condition the bush was subjected to undergo loading due to the rotation of the shaft. Two bushes, one at center and another at front end of the starter was assembled and tested for 350 starts to 1500 starts. Experimental setup is shown in fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002873_icumt.2010.5676596-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002873_icumt.2010.5676596-Figure1-1.png",
+        "caption": "Fig. 1. Coordinate frame definitions for the Quad Rotor UAV. QRUAV is modeled as a rigid combination of a central cylinder, four extending cylindrical arms and four cylinders at the end of each arm representing motors.",
+        "texts": [
+            " The dynamic model for this research closely follows the derivation of [6] and [7] which considers both gyroscopic effects and some components of drag. The resulting equations are summarized below. Let {E} be the earth fixed inertial frame, and a vector denote the position of the center of mass of the air frame as expressed in {E}. Let {B} be a body fixed frame. The orientation of {B} is defined using Q 978-1-4244-7286-4/10/$26.00 \u00a92010 IEEE Euler angles (Roll), (Pitch), and (Yaw), in that order. And is the rotational transformation matrix from {E} to {B}.Fig. 1.illustratesthe defined coordinate system. Then (1) 001 001 (2) \u2126 \u2126 \u2126 (3) where denotes the linear velocity of the airframe expressed in {E} and denotes the angular velocities of the airframe expressed in {B} [8]. Mass of the UAV is denoted as m and denotes the constant inertia matrix around the center of mass of the vehicle expressed in {B}. Here is the thrust generated by the rotor, vector is the total forces generated by the friction forces on the QRUAV structure expressed in {B}. For the purpose of this study only the aerodynamic friction resulting due to the structure of the QRUAV is considered"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000689_09544100jaero206-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000689_09544100jaero206-Figure6-1.png",
+        "caption": "Fig. 6 CTAT typical flight trajectory",
+        "texts": [
+            " It was found that the specified nozzle is capable of vectoring the engine\u2019s exhaust gases up to 29\u25e6. A sample graph of thrust deflection angle for Olympus engine in 78 000 rpm, which is found by CFD analysis and experiments, is shown in Fig. 5. A generic code, linked to the CTAT non-linear model, is used to produce the values of controls and states in every predefined trim condition. This code is based on the numerical algorithm that iteratively adjusts the independent variables until a steady state solution criterion is met. As shown in Fig. 6, the CTAT predefined flight trajectory is composed of three major parts; takeoff transition, level flight and pull-up to vertical. In order to assess the aircraft general capabilities and find out maximum required thrust-vector angle, the trim analysis is performed in some points over this trajectory. Turn analysis is also carried out so as to evaluate JAERO206 \u00a9 IMechE 2007 Proc. IMechE Vol. 221 Part G: J. Aerospace Engineering at YORK UNIV LIBRARIES on July 12, 2012pig.sagepub.comDownloaded from the CTAT lateral-directional motion and estimate the accuracy of the equations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure3-1.png",
+        "caption": "Figure 3. Geometric relation between current position and target position.",
+        "texts": [
+            " In a sense, this control scheme adopts a rolling horizon method in order to adapt to changes in road geometry in a very smooth way. A performance test of the developed control method has been carried out over a period of years, and the test results are shown in the Test Result and Analysis Section. 3. GENERATION of THE PATH PLAN The target position of a vehicle at (x, y) will be denoted as T(x, y, \u03b8) where theta indicates the difference between the current direction and the target direction, as shown in Figure 3. Since the current direction is always identical to that of the Yw axis, the difference in directions will be represented as theta, which is an angle value between the Yw axis and target direction. The target direction line intersects with the Yw axis at P, the coordinate position of which will be denoted as (0, Py) hereafter. Our objective is to drive a vehicle from the current position at (0,0,0) to the target position at (x, y, \u03b8) with a minimum change in steering angle in order to create a smooth path. As mentioned earlier, the travel path will be composed of different lines or arcs. Since a straight line can be considered as an arc with infinite curvature, a travel path from the current position to a target position will be composed of one or more connected arcs of different radii. For easy derivation of a travel path, two additional notations are introduced, l1 and l2, as indicated in Figure 3. where the length of l1 will always be equal to the Py value of point P. The target position can be situated either in the first quadrant or in the second quadrant of the world coordinate system. In this paper, only the first quadrant will be considered since the derivation in the second quadrant will be similar to that in the first quadrant. Depending upon the geometry of the target direction line, the following different cases will be considered: Case 1. (Py \u2265 0 and |\u03b8| \u2264 90o) This is the case when the y-position of P, Py is positive, and the directional difference value, theta, is between the current and the target directions is less than 90 degrees"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001763_s11666-010-9534-8-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001763_s11666-010-9534-8-Figure2-1.png",
+        "caption": "Fig. 2 High-frequency microforging apparatus. (a) Ultrasonic type. (b) Mechanical type",
+        "texts": [
+            " Therefore, high-frequency microforging can control the cracking behavior of the layer in two aspects: (1) to transform the layer organization from as-cast to as-forge, to weld microcracks, to refine microstructure of coating and (2) to form compressive residual stress to offset the tensile stress during cooling. High-frequency microforging apparatus is divided into three types: mechanical, ultrasonic, and electromagnetic. The former two were adopted in this experiment. The ultrasonic type is made up of a forging punch, an amplitude variation pole, a piezoelectric ceramic ultrasonic vibration generator, and a control power supply box (as Fig. 2a shows). It uses micrometer-scale amplitude and high-frequency vibration energy to push the forging punch to forge the surfaces, which produces an incident wave of compressive stress. Then on the interface of cladding layer and substrate, a compressive stress reflection wave forms. The superposition of the two waves forms a high compressive stress district and leads to microscale ultrathin Journal of Thermal Spray Technology Volume 20(3) March 2011\u2014457 P e e r R e v ie w e d compressive plastic deformation. The mechanical type comprises frequency modulation motor, cam, and microforging punch (see Fig. 2b). To make it convenient for analysis, suppose the propagation characteristic of stress wave was under onedimensional condition microforging only, as shown in Fig. 3(b); therefore, the equation of incident stress wave (Ref 32) produced by microforging should be: @2u @t2 \u00bc a2@ 2u @x2 \u00f0Eq 1\u00de While in medium I (coating), a1 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffi E1=q1 p ; while in medium II (substrate), a2 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffi E2=q2 p : Here E and q, respectively, stand for the modulus of elasticity and the density"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure5-1.png",
+        "caption": "FIG . 5. Hydrostatic control wheel spindle unit.",
+        "texts": [
+            " This and the fact that a loaded grinding wheel spindle will usually be tilted from the bearing axis requires that an adequate margin of safety should be allowed on the maximum applied load. However, with oil it is not diffcult to obtain the necessary load-bearing capacity as may be seen from the data in Fig. 6. If \u00df = 0-5 and y = 0-5, then Wmax = 0-55. Grinding Machine Spindles 459 The layout of the spindle was largely determined by the decision to incorporate the spindle and bearings in a cartridge assembly, Fig. 5. A hardened steel spindle was supported in a cast iron bearing housing having four pockets. The pockets would present a problem by the traditional methods of machining but the manufacture was easily and economically carried out by spark machining. A pulley and belt drive replaced the geared drive fitted with the plain bearing spindle. Oil was taken from a manifold block to each bearing pocket individually by removable stainless steel tubes. The tubes act as capillary restrictors and were calibrated for resistance before fitting"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000718_med.2008.4602251-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000718_med.2008.4602251-Figure1-1.png",
+        "caption": "Fig. 1. A three-wheeled (two driving, one castor) mobile robot.",
+        "texts": [
+            " The paper is organized as follows: in section II the model of the mobile robot is presented; in section III we shortly review the Extended Kalman Filter and in section IV we give a simple probabilistic model for the missing observations problem. In section V we propose the scanning technique to improve the accuracy of the state estimate and finally in section VI we present and discuss the results of the numerical simulations. The mobile robot we take into consideration has two driving wheels and a castor wheel, as shown in Figure 1. The robot motion in the x-y plane may be described by the discretetime model [1], [2]      xk+1 = xk + vk T cos(\u03b8k+1) + wl k yk+1 = yk + vk T sin(\u03b8k+1) + wr k \u03b8k+1 = \u03b8k + \u2206k + w\u03b8 k 978-1-4244-2505-1/08/$20.00 \u00a92008 IEEE 1850 where: \u2022 xk and yk denote the position of the center of the robot axle at time k \u2022 \u03b8k is the angle between the robot axle and the x-axis at the same instant \u2022 T is the sampling period \u2022 vk = R (\u03c9l k + \u03c9r k)/2 is the linear velocity of the robot, where \u03c9l k and \u03c9r k are the angular velocities of the left and the right wheel, respectively, and R is the radius of the wheels \u2022 \u2206k = R (\u03c9l k\u2212\u03c9r k)/l is the rotation within the sampling period, where l is the length of the robot axes \u2022 wl k, wr k, w\u03b8 k are Gaussian noises, which are assumed zero-mean and uncorrelated each other; see [2] for a real-life mobile robot modeled by the above equations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000070_978-1-4020-4941-5_42-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000070_978-1-4020-4941-5_42-Figure5-1.png",
+        "caption": "Figure 5. Over-constrained mechanisms.",
+        "texts": [
+            ", define a linear variety of lines with rank 3, i.e. a plane of lines defined by k0 and i1. Family-2 (Fig. 3): The axes of the legs in the set 1 1 1 IQ Q define a linear variety with rank 2, i.e. a planar pencil of lines with center in R. Vertechy and V. Parenti-Castelli 390 Moreover, addition of legs of type-1 and/or of type-2 to such basic USPMs does not alter the mechanism kinematics but renders the systems redundant and with self-motion. Examples of over-constrained US-PM with five and six US-legs are depicted in Fig. 5. For ease of understanding, the redundant legs are drawn in long-dash-dot lines. Note that over-constrained architectures have several advantages with respect to the basic ones. Indeed, the former make it possible to augment the mechanism stiffness-to-weight and -to-encumbrance ratios, the mechanism strength-to-weight and -toencumbrance ratios, allow the mechanism to be preloaded so as to reduce system backlash, and allow the system to be built through simpler elements such as rafters and wireropes. As an example, the mechanism depicted in Fig. 5.a can be made by means of one rafter (leg drawn in long-dash-dot line) and by four wireropes (legs drawn in solid line). Besides, over-constrained architectures further limit the range of motion of the mechanism and render its assemblage more complex. From a kinematic standpoint, note that many of the U and S joints of the mechanisms depicted in Figs. 2-5 may, in practice, be suppressed and/or replaced by simpler pairs. Indeed, joints which are not placed along the axes k0 and i1 are idle; the joints which are placed either on k0 or i1 work as simpler revolute joint with rotation axis along k0 or i1, respectively; and the joints which are placed on both k0 and i1 work as U joints with rotation axes along k0 and i1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000521_s11012-006-9049-z-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000521_s11012-006-9049-z-Figure2-1.png",
+        "caption": "Fig. 2 (a) Photograph of the spinning wheel with universal joint, and handle containing the load-cell, plus exterior highgain amplifier. The motor for spinning-up the wheel is also shown, top left. (b) Schematic diagram of the spinning wheel plus load-cell system to measure its weight",
+        "texts": [
+            " (3) Although this analysis has been straightforward, the weight-reduction phenomenon appears unacceptable to those who cannot visualise such experimental techniques; yet no alternative energy equation or discussion has ever been forthcoming from them. If weight-reduction did not occur because Eq. (1) were wrong, then the horizontal input impulse energy would have disappeared somehow. In practice, it is difficult to input the correct amount of impulse energy, so an equivalent technique demonstrated by Laithwaite has been pursued, as follows. A photograph of the experimental apparatus is shown in Fig. 2a, and the apparatus schematic diagram is shown in Fig. 2b. A thick-rimmed steel wheel of mass M = 3 kg and mass-weighted outer diameter 182 mm has a moment of inertia I = 0.025 kgm2. It is mounted with ball bearings onto an axle of length = 260 mm, and may be spun up to 2500 revolutions per minute by means of a detachable motor. Then under the influence of gravity, the normal precession velocity is: vp = Mg 2 I\u03c9 = 0.31ms\u22121 . (4) The other end of this dural axle is attached to a universal-joint which allows up to 60\u25e6 of misalignment from its axis in any direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003536_mcs.2011.941250-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003536_mcs.2011.941250-Figure1-1.png",
+        "caption": "FIGURE 1 Aircraft and Earth frames. The aircraft frame is fixed to the aircraft, while the Earth frame is assumed to be an inertial frame. The signed quantities , and h determine the location of the point p at which the output is defined relative to the center of mass c. The pitch angle Q, which is positive as shown, is determined by the right-hand rule about the axis j\u0302AC 5 j\u0302E, which is not shown but is directed out of the page.",
+        "texts": [
+            " AIRCRAFT KINEMATICS The Earth frame FE, whose orthogonal axes are labeled i\u0302E, j\u0302E, and k\u0302E, is assumed to be an inertial frame, that is, a frame with respect to which Newton\u2019s second law is valid [19]. A hat denotes a dimensionless unit-length physical vector. The origin OE of the Earth frame is any convenient point on the Earth. The axes i\u0302E and j\u0302E are horizontal, while the axis k\u0302E points downward; we assume the Earth is flat. The aircraft frame FAC, whose axes are labeled i\u0302AC, j\u0302AC, and k\u0302AC, is fixed to the aircraft. The center of mass c and frame vectors i\u0302AC and k\u0302AC are shown in Figure 1. The aircraft is assumed to be a three-dimensional rigid body. In longitudinal flight, the aircraft moves in an inertially nonrotating vertical plane by translating along i\u0302AC and k\u0302AC and by rotating about j\u0302AC. The direction of j\u0302AC is thus fixed with respect to FE. For convenience, we assume that j\u0302AC 5 j\u0302E. The velocity and acceleration of the aircraft along j\u0302AC are thus identically zero for longitudinal flight, as are the roll and yaw components of the angular velocity of the aircraft relative to the Earth frame. The sign of the pitch angle Q, which is the angle from i\u0302E to i\u0302AC, is determined by the right-hand rule with the thumb pointing along j\u0302AC and with the fingers curled around j\u0302AC. For example, the pitch angle Q, shown in Figure 1, is positive. Let p denote a point in the plane that is parallel to the i\u0302AC - k\u0302AC plane and passes through c. The position of p relative to OE can be written as rYp/OE 5 rphi\u0302E 1 rpvk\u0302E, (1) where a harpoon denotes a physical vector. The position of p relative to c is given by rYp/c 5 rYp/OAC 1 rYOAC/c 5 rY p/OAC 2 rYc/OAC , (2) which can be written as rYp/c 5 ,i\u0302AC 1 hk\u0302AC, (3) where , . 0 indicates that p is forward of c, that is, toward the nose, and , , 0 denotes that p is aft of c, that is, toward the tail"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003133_1350650112464323-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003133_1350650112464323-Figure4-1.png",
+        "caption": "Figure 4. Geometric parameters and design variables (ts and ).",
+        "texts": [
+            " To reduce the execution time, this number can be adaptively increased as the deformation becomes close to its equilibrium value. A typical execution time was roughly 20 minutes on a server computing environment equipped with an Intel Xeon processor E5630, and 16GB of memory. The objective of the optimization is to maximize the LCC of a bearing by finding the optimum dimensions of the split. The design variables are the pad thickness above the split (ts), and split angle ratio ( \u00bc s= 0) as shown in Figure 4. Generally, the split height (hg) is not an important factor as long as it allows the deformation of the top surface without contacting the lower surface. As a result, this parameter is considered to be fixed at 2mm in this study. The optimization method used here is based on a hybrid of the harmony search (HS) algorithm and sequential quadratic programming (SQP) method. at Monash University on December 6, 2014pij.sagepub.comDownloaded from There is a vast literature on both algorithms so the details are not repeated here"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure2.38-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure2.38-1.png",
+        "caption": "Figure 2.38 Three-dimensional interpretation of the Nichols closed loop gain curves",
+        "texts": [
+            " There can be some difficulty for the new user of this tool to overcome the visual complexity associated with all of the closed loop gain and phase curves superimposed on the open loop coordinates. This can sometimes make it difficult to observe what is really going on. One technique, mentioned briefly earlier, that may be helpful to the control systems engineer is to focus on the closed loop gain curves of the Nichols chart and to visualize these gain values as a third dimension coming out the chart as interpreted by the artistic impression presented as Figure 2.38. As shown the closed loop gain curves become progressively higher as they approach the instability point This \u2018mountain\u2019 rising out of the page implies that a response curve passing close to the instability point (0 dB, \u2212180 degrees open loop), as is the case when stability margins are small, will have to cross the \u2018mountain\u2019 at a \u2018high altitude\u2019 resulting in the closed loop gain response being greatly magnified for the frequencies corresponding to the high points on the \u2018mountain\u2019. This effect is equivalent to the resonance displayed by the spring\u2013mass system analyzed in the first chapter"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001956_1.3601566-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001956_1.3601566-Figure1-1.png",
+        "caption": "Fig. 1(b) Face seal, noninert ia l coordinate system",
+        "texts": [
+            " This paper is concerned with: (a) The application of the theory, including inertial effects, to those situations where the seal geometry deviates from the ideal aligned, flat surface, and an assessment of the relative effect of these deviations. (b) The special hydrodynamic features of the flow in the clear- ance space which are directly attributable to inertial effects, e.g., separated flow and cavitation. (c) An estimation of the limitations of the laminar theory, i.e., the maximum value of the Reynolds number for which the theory is valid. Analysis of M i s a l i g n e d Face Seal Film Thickness Equation. The geometry of the seal is shown in Fig. 1(a). I t will be convenient to adopt a coordinate system rotating with the precessional angular velocity 0 , since in this system the time-dependence at a fixed coordinate position is eliminated. In Fig. 1 (b ) the x axis is chosen so that it always passes through the point of minimum film thickness. With this axis as a reference line for a rotating cylindrical coordinate system, the equation for the film thickness is given b y 11 = 110 \u2014 r cos 6 tan X (1) Equations of Motion. The diniensionless variables of the system will be defined bj r the following equations: 1 Numbers in brackets designate References at end of paper. Contributed by the Lubrication Division and presented at the Winter Annual Meeting, Pittsburgh, Pa"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003982_clen.201200075-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003982_clen.201200075-Figure2-1.png",
+        "caption": "Figure 2. COD removal efficiencies of the reverse osmosis membranes.",
+        "texts": [
+            " As for the average permeation flux obtained with the ultrafiltration process, it was determined to be 34.0 L m 2 h 1. Reverse osmosis studies were carried out using two separate approaches for each membrane: one following the centrifuging process and another for the centrifuging\u00feultrafiltration process. Therefore, 16 samples of filtrate water were obtained from both membranes under four different levels of pressure. The COD concentrations measured in these filtrate waters and the obtained COD removal efficiencies are shown in Fig. 2. From Fig. 2, it was observed that with increasing transmembrane pressures, the COD removal efficiencies were also increased. This increase is in a linear relationship to the transmembrane pressure. The COD removal efficiencies of the samples pre-treated by means of centrifuging only were found to be somewhat higher. While wastewater passes through the membrane system, the particles are retained on the surface of the membrane due to them being larger in size than the membrane diameter. In time, these particles may cause the membrane pore size to reduce"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001077_detc2009-87092-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001077_detc2009-87092-Figure3-1.png",
+        "caption": "Figure 3. FE model of the cycloid drive.",
+        "texts": [
+            " 1 2 (10ln 2 ) 1 1p bEk \u03c0 \u03bd \u03bd \u03bd \u2212= \u2212 \u2212 \u2212 (2) The cycloid disk is fixed at its center, and no external force is applied to the disk. On the other hand, two couple forces for rated torque are applied only to the housing whose rotational displacement around its own center is allowed. Contact forces from FE analysis are in good agreement with theoretical calculations, as shown in Figure 2. The maximum error between FE analysis and theoretical calculation is less than 2%. A FE model for a two-stage cycloid drive is shown Figure 3. A rolling element bearing between the eccentric shafts and the cycloid disk is modeled as a rigid ring that has frictional contact (\u03bc = 0.001) with the disk and is elastically supported, as shown in Figure 3. By using a simple FE model of the bearing, rigidity of the elastic support is adjusted to attain theoretically calculated bearing rigidity [6]. Calculated bearing rigidity is slightly nonlinear and amounts to 2.12 \u00d7 108 N/mm at the rated torque of the cycloid drive. Because taper and needle roller bearings are serially connected through the eccentric shaft, bearing stiffness decreases by half. Corresponding rigidity of the elastic support element is set to 2.2 N/mm3 at rated torque. Because the cycloid disk is expected to move due to finite bearing stiffness and contact force distribution will change significantly, spring elements for all 40 contact points are included initially, and spring elements that are supposed to have tensile forces are removed through iterative FE analysis"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003552_j.bios.2013.02.046-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003552_j.bios.2013.02.046-Figure2-1.png",
+        "caption": "Fig. 2. Device design and SEED measurement. (a) Schematic showing the two main steps to immobilize GOx on the gradient electrode. (b) Schematic of the top view of the sample chamber confined within the \u201cO\u201d ring. The \u201ctwo (red) dots\u201d are laser spots. The inset at the top is the optical microscope image of the electrode showing the \u201cV\u201d pattern formed by the resist. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article).",
+        "texts": [
+            " The maximum redox current is linearly proportional to the observed \u0394max (Fig. 1b). Thus, measurement of local \u0394max quantitatively maps the relative redox current density. The \u0394max is an average (of the oxidation peak) over five cycles. The error bars averaged over five V-ramp cycles are less than 0.4 pm (OSD, Fig. B). The gold (Au) electrode is coated with 5 and 50 mm polymer photoresist SU-8, exposed to ultraviolet (UV) light, and developed to form a \u201cV-shaped\u201d surface exposed to the solution (Fig. 2a, Step 1). The resist acts as the passivation layer (similar to PMMA in Fig. 1). The exposed \u201cV-shaped\u201d Au electrode, similar to Fig. 2b inset, is coated with three alternating \u223c0.9 nm thick layers of poly (allyl amine hydrogen chloride) (PAH) (15,000 Da, Sigma Aldrich,) and poly(styrene sulfonate) (PSS) (70,000 Da, Sigma Aldrich) (Steps 1 and 2, Fig. 2a) for enzyme immobilization. The PAH and PSS films are spin coated at 3000 RPM from aqueous solutions of 3 mg/ml each. The glucose oxidase (GOx) enzyme dissolved in 100 mM 2-(N-morpholino)ethanesulfonic acid (MES) buffer (pH 5.7) is placed on the sample and left covered overnight in a refrigerated environment to form an immobilized monolayer (Step 2, Fig. 2a). Although the whole surface after PAH/PSS coating is hydrophilic, due to the etched grove in SU8, the enzyme solution drop is confined over the electrode. The requirement of confinement of enzyme on the electrode is not essential because deposition of enzyme over SU-8 will not matter as no redox can occurs over the passivated region. The enzyme is encapsulated with another layer each of PAH and PSS, and the measurements are performed with laser beams on the sample and resist (Fig. 2a). A designed electrode was placed in a sample chamber, as shown in Fig. 2b. One beam is placed on SU-8 film working as a passivation layer, while the other beam is measuring active electrode reaction. The change in local Imax on a monolith electrode is measured by scanning the laser beam. To note is that lithography is a convenient method for making a patterned monolith electrode for SEED where the resist has to be stripped, which requires environmentally caustic chemicals. The local size of the electrode is defined by p in the optical microscope image (Fig. 2b, inset). The apex angle is \u223c251, and the maximum p (not seen in the microscope image) is \u223c1 mm. The dielectric resist will focus the field due to dielectrophoresis causing larger current density as p decreases (Christensen and Hamnett, 2000). The laser (schematically shown as a \u201cred dot\u201d) is scanned from low to high p; and the optical signal from local redox of 50 mM [Fe(CN)6]4\u2212/3, similar to Fig. 1b, is recorded. Consistent with the dielectrophoresis effect, \u0394max (averaged over five scans) increases as p decreases (Fig",
+            " However, the SEED and CV signals in the oxidation cycle are different because the latter is due to H2O2 oxidation. The nonturnover SEED signal is a superposition of five cycles. The measurement of a monolayer of enzyme during redox with no turnover is an important aspect of SEED to study characteristics of an enzyme, i.e., protein voltammetry (Leger et al., 2003). Several potential applications for SEED may be envisioned. For example, a multianalyte chip can be fabricated by patterning \u223c10 mm features in SU8 resist (similar to Fig. 2a) followed by spot-by-spot immobilization of probe molecules, such as ssDNA oligomers, and different enzymes. The probe molecule solutionwill be localized over each feature by surface tension. (The surface tension effect is observed during the fabrication of devices described in Fig. 2a.) The machine for spotting can be the same as that used in microarray fabrication. In another application, as the signal is insensitive to mechanical vibrations from motion and thermal fluctuations, SEED can be a sensitive readout system for transparent microfluidics devices where the probe beam is incident in the device channel while the reference beam is outside the channel. This may lend itself to multianalyte analysis on small volumes, and multiple sites in the microfluidics device could be interrogated"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000485_j.fss.2008.03.021-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000485_j.fss.2008.03.021-Figure12-1.png",
+        "caption": "Fig. 12. Planar two-link flexible-joint manipulator arm.",
+        "texts": [
+            "1) where the weight drift has been prevented without sacrificing performance (Fig. 11). Moreover, the weight drift increases gradually as is decreased, rather than unexpectedly. This is a large advantage in qualitative behavior over leakage and e-modification. If a control designer is varying parameters in search of better performance, the system will slowly get worse instead of suddenly and unexpectedly. A two-link flexible-joint robotic arm (referred to as a 2 DOF Serial Flexible Joint by the manufacturer) serves as an experimental test bed (Fig. 12). An arrangement of adjustable linear springs provides the elasticity in the joints, allowing the experimenter to easily change the (equivalent rotational) spring stiffness. The gearing is supplied with harmonic drives with gear ratios of 100:1 and 80:1 (for rotors 1 and 2, respectively) whose flexibility results in potentially significant higher-order dynamics. The manufacturer supplies some mechanical model parameters (Table 2). A payload of 1 kg rides on the tip of the robot during the experiments"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002169_iros.2009.5354046-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002169_iros.2009.5354046-Figure8-1.png",
+        "caption": "Fig. 8. Cross section of minimal configuration",
+        "texts": [],
+        "surrounding_texts": [
+            "No propulsion\nReaction force as propulsion\nInsertion Small drag & buckling\nFig. 2. Snake-like mechanism as an endoscope; Lateral undulation does not function during less curved section. It produces propulsion at corners where a passive device might fail to progress due to buckling\n\u03c6i\n\u03c6i\nA\nB\nlink i \u2212 1\nlink i + 1\nlinki\n\u03c6i+1 \u2212 \u03c6i = 0\nFig. 3. Propagation of joint angle: Difference in two joint angles \u03c6i+1 \u2212 \u03c6i appears in the relative angle of gear B to i + 1th link\nThe proposed mechanism consists of mainly two components. One is the transfer mechanism of joint angle. In order to simulate curvature derivative, adjacent joints must share the joint angle information. The other is fluid servomechanism.\nCurvature derivative is approximated by the difference in joint angles of the posterior joint and anterior joint as expressed in (4). The following are important in mechanical implementation:\n\u2022 Zero error must be realized by mechanical agreement \u2022 Rotational difference must be realized on a common\naxis\nFig. 3 shows how the above requirements can be realized using a pair of identical gears. Mechanical elements painted\nwith the same pattern denotes mutually fixed or geared elements. Three independent groups of elements are drawn. The gear A is attached to the link i\u22121 and geared to the gear B. An arrow mark is drawn on each gear and link to indicate their original position. From the left figure to the middle, the i \u2212 1th joint is bent clockwise at angle \u03c6i. Then the gear B turns 2\u03c6i clockwise. So the relative angle between i + 1th link and the gear B is identical to \u03c6i. In the right figure, the i+1th link is bent at the same angle \u03c6i clockwise relative to the ith link, where both arrow marks are congruent. Hence, the difference of joint angle \u03c6i+1 \u2212\u03c6i can be formed as the angle between the two arrow marks.\nConventional hydraulic power steering system for automotive can be regarded as a servomechanism where the steering wheel angle is the reference angle and the actual",
+            "wheel angle is the output. Valve mechanism is linked to the feedback error, namely, the difference between the input and output. Thus hydraulic power is supplied to attenuate the error as long as the valve is open. Fig. 4 shows the basic action of a standard power steering system. The size of valve orifice depends on the relative position of the rotary and outer valves. Thus the steering angle can be stabilized at arbitrary angle.\nC. Integration\nFluid servomechanism must have three independent elements; input, output, and base. These three elements should be constructed on each joint. On the ith joint, the input is connected to the gear B. The output is connected to the i+1th link. The base in this case is the ith link. The actuator is located between ith link and i + 1th link to produce torque between these elements. The valve system is organized to be dependent on the relative position of the input and output. When there is a difference in angle between the gear B and i + 1th link, a gap appears in the valve to drive the actuator so that the gap becomes smaller.\nThe first unit determines the direction of progress because the posterior units follow the same joint angles afterwords. The operator manipulates this part like a steering of a vehicle via wire or other remote control methods as shown in Fig. 5. If there exists a wall that can guide the head unit, the guide rollers will suffice for exploration.\nFig. 6 shows the front view of the prototype. The maximum bending angle of each joint is about 14 degrees. The center distance of each joint is 16 (mm), so the minimal inner radius is approximately 100 (mm). The head unit has guide rollers on the both sides, so no wire is installed to manipulate it. The materials consists of aluminum alloy"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002478_dscc2012-movic2012-8641-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002478_dscc2012-movic2012-8641-Figure4-1.png",
+        "caption": "Figure 4. TURNING MANEUVER OF SKID-STEERED VEHICLE ON HILL.",
+        "texts": [
+            "url=/data/conferences/asmep/76066/ on 04/04/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use the vehicle. Following [3], the inner and outer sliding velocities are given by Vji = (Vjxi,Vjyi), (17) where Vjxi =\u2212yi\u2126z and Vjyi = (R\u2032\u2032\u2212 B 2 \u2212Cx +xi)\u2126z\u2212 rwi. Similarly, Vjo = (Vjxo,Vjyo), (18) where Vjxo =\u2212yo\u2126z and Vjyo = (R\u2032\u2032\u2212Cx + B 2 + xo)\u2126z\u2212 rwo. Then, \u03b3i and \u03b3o can be computed as \u03b3i = arctan( Vjyi Vjxi ), (19) \u03b3o = arctan( Vjyo Vjxo ). (20) The normal pressures on the wheels vary as a function of the terrain slope and the vehicle heading. Referring to Fig. 4, it is possible to perform a moment balance about the \u201cinner and outer wheel lines\u201d (i.e., the line connecting the center of the inner wheels and the corresponding line for the outer wheels) and obtain the following normal forces on the inner (Ni) and outer (No) sides as Ni = W cos\u03b8(B 2 \u2212Cx)+Whsin\u03b8sin\u03c8\u2212 Whv2 gR\u2032 cos\u03b2 B , (21) No = W cos\u03b8(B 2 +Cx)\u2212Whsin\u03b8sin\u03c8+ Whv2 gR\u2032 cos\u03b2 B , (22) where \u03b8 and \u03c8 represent the slope of the hill and the vehicle heading, W is the vehicle weight, which as shown in Fig. 4 has a longitudinal component Wln and a lateral component Wlt , h is the height of the CG, R\u2032 is the distance from O to the CG, and \u03b2 represents the angle between the vector going from the center of turn O to the CG and a vector going from O to Ov (refer to Fig. 3 for easy visualization of this parameter). Similarly, performing a moment balance about the front and rear axles, we get N f = W cos\u03b8( l 2 +Cy)\u2212Whsin\u03b8cos\u03c8\u2212 Whv2 gR\u2032 sin\u03b2 l , (23) Nr = W cos\u03b8( l 2 \u2212Cy)+Whsin\u03b8cos\u03c8+ Whv2 gR\u2032 sin\u03b2 l , (24) Assuming symmetry, the normal pressures on each wheel are then estimated using po f = N f W cos\u03b8 No b pl , (25) por = Nr W cos\u03b8 No b pl , (26) pi f = N f W cos\u03b8 Ni b pl , (27) pir = Nr W cos\u03b8 Ni b pl ",
+            " Once the rolling resistances are determined, it is possible to estimate the coefficients of rolling resistances for the inner and outer sides as follows: \u00b5ri = Ri Ni , (29) \u00b5ro = Ro No . (30) 5 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76066/ on 04/04/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Finally, the total resistance term C(q, q\u0307) can be expressed by C(q, q\u0307) = r [ Fi +Ri Fo +Ro ] + [ \u03c4ires \u03c4ores ] , (31) where r is the wheel radius. When traversing slopes, it is also necessary to overcome the gravitational term G(q), which can be derived from Fig. 4 by performing a force and moment balance as G(q) = rW sin\u03b8cos\u03c8 B [ B 2 \u2212Cx, B 2 +Cx ]T . (32) As shown in Fig. 2, when non ideal actuators are pushed to their saturation limits, uncontrollable vehicle motion could result. These situations can be avoided by incorporating motor and motor controller models into the vehicle dynamic model. In this research, the modified Pioneer 3-AT robotic platform uses two Maxon motor controllers (one for each vehicle side) controlled in current mode. In order to protect the motors, maximum conservative current levels of Ilim = 5",
+            " 7 for the outer and inner sides are 0.2186N and 0.1362N respectively. This section presents the experimental validation of the model on a slope employing the same wood platform used for the estimation of the terrain dependent parameters. First, experimental results are presented for linear motion on the hill with an arbitrarly heading. Second, we evaluate the model performance under steady circular motion. In all experiments, the heading angle \u03c8 is measured positive counter clockwise as illustrated in Fig. 4. Figure 8, presents the torque results for the vehicle moving straight with an initial heading of 60\u25e6 and a hill inclination of 10\u25e6. The corresponding RMSE for the inner and outer sides are 0.1355Nm and 0.1445Nm respectively. It is important to clarify that the data corresponding to the first 0.2s was excluded in the calculation of the errors since this data corresponds to the initial transient and as mention in Sec. 2, the model was developed under steady state assumptions. Fig. 9 corresponds to torque of the vehicle moving straight with an initial heading of 0\u25e6 and a hill inclination of 15\u25e6"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001854_gt2009-59580-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001854_gt2009-59580-Figure3-1.png",
+        "caption": "Figure 3. Bladed Disk example: (a) view, (b) blade sector finite element mesh, and (c) natural frequencies and coupling indices vs. nodal diameter plot.",
+        "texts": [
+            " In these results, n denotes the size the matrix iiY ( ) 12 \u2212\u03bb+\u2212= inip A , 2 12 \u2212\u03bb+ =\u03bc n , and \u03bb>0 represents the single free parameter of the stochastic model. This parameter can be evaluated to meet any given information about the variability, e.g. standard deviation of appropriate natural frequencies, see the example section for discussion. (figures for n=8, i=2, and \u03bb=1 and 10). nloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Term The above discussion was exemplified on the bladed disk model shown in Fig. 3 which is a modification (a reduction of the number of blades to 12) of the one considered in [5]. Since the original model is a blisk and thus does not exhibit a detailed modeling of the blade root as it fits in the disk, the simple interface modeling of Eq. (13) and (14) was adopted. The blade-interface-disk modeling was extracted directly from the sector finite element model using the cyclosymmetric CraigBampton formulation of [5]. The frequency vs. nodal diameter plot of Fig. 3(c) above has been enriched with a measure of blade to disk coupling recently introduced [22] and referred to as the coupling index (abbreviated by ci). This measure of coupling relates to the change in the cyclic system natural frequency corresponding to a particular number of nodal diameters induced by a change of blade only (not disk) Young\u2019s modulus. If the resonance condition investigated is associated with a blade alone motion, then a relative change E\u03b4 of blade Young\u2019s modulus would affect the blade alone natural frequencies by the factor E\u03b4+1 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000586_j.bios.2008.07.084-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000586_j.bios.2008.07.084-Figure1-1.png",
+        "caption": "Fig. 1. Scheme of the flow system for the trehalose determination (1: sample solution; 2: peristaltic pump; 3: trehalase reactor; 4: GOD biosensor in the flow cell; 5: amperometer; 6: recorder; 7: switch (T valve); 8: waste. (A) Path to assay the total glucose in the sample (the intrinsic one plus the one produced by the trehalase); (B) path to assay the intrinsic glucose in the sample. Inset: [ - (1) to",
+        "texts": [
+            " Apparatus Amperometric measurements were carried out using a VA 41 amperometric detector (Metrohm, Herisau, Switzerland), conected with a X-t recorder (L250E, Linseis, Selb, Germany). Cyclic oltammetry experiments were performed with an Autolab elecrochemical system (Eco-Chemie, Utrecht, Netherlands) equipped ith PG Stat12 and GPES software. The orbital shaker was from PM Instruments (Bernareggio, Milan, Italy). Peristaltic pumps ere Minipuls3 (Gilson, France), Rotavapor (B\u00fcchi, Labortechnik G, SW). The trehalase reactor was a home-made Teflon apparatus, and drawing is shown in Fig. 1. The electrochemical cell of the \u201cthinayer\u201d type was a plexiglass home-made apparatus, formed by a ow-through cell with two tubes (in and out) and the appropriate older for the GOD electrode (Fig. 1). .3. Electrodes Screen-printed electrodes (SPEs) were home-produced with a 45 DEK (Weymouth, England) screen-printing machine. Graphiteased ink (Elettrodag 421) from Acheson (Milan, Italy) was used o print the working and counter electrode. The substrate was a exible polyester film (Autostat HT5) obtained from Autotype Italia Milan, Italy). The electrodes were produced in foils of 48 strips. Each sensor onsists of three screen-printed elements: two carbon electrodes, orking and counter, and a silver electrode acting as pseudorefrence, respectively",
+            "5%, v/v) glutaraldehyde solution nder stirring for 40 min. On each side of the so treated memrane, rewashed with distilled water, 13 L of trehalase solution 3.7 IU mL\u22121) were deposed and left to dry for 1 h. After rewashing, he membrane was dipped in a glycine solution (0.5 mol L\u22121) for 0 min with stirring. For preservation, the so prepared membrane as maintained in a glycerol solution (50% (v/v) in buffer pH 6.5) t +4 \u25e6C. .7. Assembling of the flow system The system devised out for the trehalose assay (Fig. 1) consists of peristaltic pump, in order to flow the solutions into the measuring ystem, a reactor containing the immobilized trehalase enzyme, a OD amperometric biosensor for the glucose assay, an amperomter and a recorder. A switch (T-valve) allows the solution to flow r into the reactor and then through the GOD biosensor (path A), r directly into the GOD biosensor (path B). By using path B, it is ossible to measure the current response due to the glucose conentration initially present in the sample solution, while through 1384 M"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001502_cdc.2008.4739357-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001502_cdc.2008.4739357-Figure6-1.png",
+        "caption": "Fig. 6. The geometric interpretation behind the algorithm used for the 2 point case.",
+        "texts": [
+            " The idea is two perform two time a \u201csingle point\u201d iteration. Such an idea it is very simple and has guaranteed results both for the 2-point cases and for the generalized one and will be described in the next subsection. A second heuristic here proposed is built in the following way. Let us build a triangle with one vertex on q1 and the other two on the circle of center q1 and radius 1 2 a\u0304Vv and whose angle on q1 is equal to \u03b8max, such that it is included in the angle \u03b80\u030212 and such that it does not intersect the segment p0q2 as depicted in Fig. 6. If we fix the two points determined by the intersection of such a triangle and the circle as the take-off point and the rendezvous one the total cost of the strategy will be teu = \u221a( 1 2 a\u0304Vv )2 + d2 1 \u2212 (a\u0304Vv) d1 cos (\u03b81) Vc + a\u0304+ + ( 1 2 a\u0304Vv )2 + d2 2 \u2212 (a\u0304Vv) d2 cos ( \u03b80\u030212 \u2212 \u03b8max \u2212 \u03b81 ) Vc + \u2212a\u0304 ( Vc + Vv 2Vc ) + a\u0304 ( Vc + Vv 2Vv ) where \u03b81 \u2208 [ 0, \u03b80\u030212 \u2212 \u03b8max ] is the one defined in Fig. 6. In order to determine an optimal \u03b81, a simple numerical optimization can be performed, as in Fig. 7. It is possible to prove that it always exist at least a choice of \u03b81 \u2208 [ \u03b80\u030212 \u2212 \u03b8max ] such that this solution, in the case Proposition 1 & 2 doesn\u2019t apply, it is always lower then the upper bound (12). For this purpose let us consider the particular solution \u03b81 = [ \u03b80\u030212 \u2212 \u03b8max ] /2 as depicted in Figure 6, where, by construction, the angle denoted as \u03b1 is always grater than \u03c0/2 Then let us draw the lines p0pto1 and q2pl1 until they touch the height segment of the triangle. Let us denote by c1 and c2 such an intersection point and by qt the center of the basis of the triangle. Because \u03b1 > \u03c0/2 then, from the properties of the triangles, \u2016c1 \u2212 p0\u20162 < \u2016q1 \u2212 p0\u20162 and \u2016cc \u2212 p1\u20162 < \u2016q2 \u2212 q1\u20162. Let us concentrate on \u2016c1 \u2212 p0\u20162 < \u2016q1 \u2212 p0\u20162 (the same will be on the other inequalities, by symmetry). We observe that \u2016pto1 \u2212 p0\u20162 + \u2016c1 \u2212 pto1\u20162 < \u2016q1 \u2212 p0\u20162"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003969_096034012x13269868036418-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003969_096034012x13269868036418-Figure1-1.png",
+        "caption": "Figure 1 LCF specimen geometry.",
+        "texts": [
+            " With a view to assessing the influence of re-solutionising treatment on the low cycle fatigue behaviour of the alloy, the service-exposed material was subjected to a re-solutionising treatment at 1433 Ky2 h prior to testing. Cylindrical samples of 5 mm gauge diameter and 18 mm gauge length were used for the tests, which were carried out under stress control with an R-ratio of zero and a frequency of 0.1 Hz, at room temperature and 873 K. The standard code of practice outlined in ASTM E466 was used for performing the tests. Various peak stresses in the range 357 \u2013 612 MPa were employed for the tests that were performed on the virgin and service-exposed material after the re-solutionising treatment. Figure 1 presents the specimen geometry used for the tests. Metallography was carried out on samples electro-etched in saturated oxalic acid solution at 3 \u2013 5 V, using a stainless steel cathode. Table 1 presents the chemical compositions of the virgin and service-exposed materials and Table 2 summarises the results obtained from LCF tests. The prior microstructure exhibited by the virgin (V) material is presented in Figures 2(a) and (b). Figures 2(c) and (d) present the microstructures in service-exposed (SE) and service-exposed\u00fe re-solutionised (SE\u00feRS) conditions respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000724_s1064230707040077-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000724_s1064230707040077-Figure8-1.png",
+        "caption": "Fig. 8. The fragment of the model of the lorry: the rear suspension.",
+        "texts": [
+            " Unlike the approximate Jacobian matrices, for the block diagonal Jacobian matrix, the mean value of the spectral radius of the system matrix calculated on the entire integration interval grows when the stiffness of equations grows for different values of the local error, as shown in Fig. 7. We consider the results of applying the developed methods to studying complex technical systems that can be represented by the models of a lorry or a carriage. We obtained the given results using the Universal Mechanism software platform (www.umlab.ru) that allows us to study the dynamics of vehicles and rail carriages numerically. The model of the lorry consists of a detailed description of the rear suspension (Fig. 8) and a simplified description of the front suspension, contains 22 bodies, and has 39 degrees of freedom. The stiffness of motion equations is due to the elastic joints in the suspension system with large values of stiffness coefficients. We consider the results of the steering wheel hitch test, which implied the fast rotation of the steering wheel that causes the front wheels to be rotated by 10 degrees. The car moves at 100 km per hour. The simulation lasted for two seconds. When we used the approximate Jacobian matrices, the integration by the Park method with the local error \u03b5 = 10\u20136 required 3,300 calculations of motion equations and four seconds of the processor time (dual-core 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003455_iet-gtd.2010.0584-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003455_iet-gtd.2010.0584-Figure5-1.png",
+        "caption": "Fig. 5 Single-axis circuit model of a synchronous generator",
+        "texts": [
+            " After ignoring the second term, the q-axis voltage equation can be simplified to vqs = \u2212rs \u2212 p vb Xq ( ) iqs + vr vb Xmd ( ) I \u2032fd (10) For the (8), it is usually correct for Lmq \u00bc Lmd with regard to round rotor-type generators. The second term can thus be removed, and the electromagnetic torque formula can be simplified to tEM = 3 2 P 2 LmdI \u2032fdiqs = 3 2 P 2 Xmd vb I \u2032fdiqs (11) It can be seen from the simplified d-axis and q-axis voltage equations that this forms a single-axis linear equivalent circuit as shown in Fig. 5, with the electromagnetic torque formula of (11). IET Gener. Transm. Distrib., 2011, Vol. 5, Iss. 7, pp. 685\u2013693 doi: 10.1049/iet-gtd.2010.0584 The coupling between electrical and mechanical parts of a synchronous generator is determined by the back EMF and electromagnetic torque shown below e = vr vb XmdI \u2032fd tEM = 3 2 P 2 Xmd vb I \u2032fdiqs \u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 (12) By defining VB and IB as the voltage and current base values for the d\u2013q axis model, the impedance base value can be calculated as ZB \u00bc (VB/IB) and the torque base value as TB \u00bc (3/2)(P/2)(VB/IB/vb)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001502_cdc.2008.4739357-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001502_cdc.2008.4739357-Figure5-1.png",
+        "caption": "Fig. 5. The geometric interpretation behind Proposition 2.",
+        "texts": [
+            " This inequality holds true if cos ( \u03b80\u030212 ) \u2264 1 \u2212 2(V 2 c /V 2 v ) that for \u03b80\u030212 \u2208 [0, \u03c0) is \u03b80\u030212 \u2264 acos ( 1 \u2212 2V 2 c /V 2 v ) 2 Proposition 2: If the straight line between p0 and q2 touches in 2 points, respectively c1 and c2, the circle of center q1 and radius Vv a\u0304/2 and if the smallest angle defined by the segments c1q1 and q1c2 is bigger or equal than acos ( 1 \u2212 2 V 2 c V 2 v ) then it exists at least one solution such that the total cost is equal to lower bound (14) . Proof. The carrier will go directly to the last point through a straight line. By hypothesis, such a straight line will intersect a triangle, built like the one in Fig. 5, with angle \u03b8max = acos ( 1 \u2212 2 V 2 c V 2 v ) . Following the proof of Proposition 1, it is known that the time used by the carrier to follow the basis of that triangle is equal to the time the fast vehicle uses to follow the other two sides. Then using those intersection point as take-off and rendezvous point, a feasible solution of cost (14) is reached. 2 For the case not covered by the two proposition above some heuristics have been developed. A first heuristic can be simply obtained by using the 1- point solution"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001219_s12541-009-0102-4-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001219_s12541-009-0102-4-Figure7-1.png",
+        "caption": "Fig. 7 Load application points and displacement measurement points for estimating the normal, lateral, and axial compliances of the Y-axis feed system",
+        "texts": [],
+        "surrounding_texts": [
+            "A virtual prototype of the ultra-precision machine for machining large-surface micro-features, which was constructed based on ANSYS software to estimate its compliances, is presented in Fig. 10. The virtual prototype was composed of 134,729 nodes, 534,911 solid elements, and 408 matrix elements. The matrix elements were introduced in order to represent the normal and lateral stiffnesses of the hydrostatic guideways, the radial and thrust stiffnesses of the hydrostatic bearings, and the axial stiffness of the linear motors. As the boundary condition for the structural analysis and measurement, the movement of the ultra-precision machine was restricted in the vertical direction at the four supporting points of the bed, as shown by the red arrows in Fig. 10. The bed, column, cross beam, and feed tables were made of cast iron (GC300), the C-axis shaft and bracket were made of steel (SCM440, SS400), and the mover/rotor and stator of motors were made of Fe-Si. Table 3 shows the material properties used for the virtual prototype."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001343_ichr.2007.4813886-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001343_ichr.2007.4813886-Figure3-1.png",
+        "caption": "Fig. 3. Double pendulum used for the comparaison of the discretization method",
+        "texts": [
+            " This paper deals with optimization applied to 2- D robot, therefore we adapt the Newton-Euler method for computing the joint torques of a 2-D robots. For humanoid robots, the computation of the Zero Moment point (ZMP) give information about the balance. [6] defines ZMP as the point zmp, on the contact surface, where the moment is equal to zero Mzmp = 0 (cf. fig 2). If this point stays in the base of support, the robot maintains its equilibrium. The zmp location depends on the joint angle q(t), velocity q\u0307(t) and acceleration q\u0308(t). The double pendulum (cf. fig.3) is used, in section IV, as a simple model. We define the initial value qi = [0,0] to get the foot position : (x = 0, y = 0). The final joints value is computed to do a step lenght of d : q f = [ Acos ( d 2L ) ,2\u00d7Acos ( d 2L )] (12) Let us consider one parameter per joint. So, in the case of Fig.3, we get 2 parameters : P = [p1; p2]. The goal is to determine the best value of P, which minimizes the energy consumption and guarantees the contraints shown in I-A, thanks to the constrained optimization algorithm. We define the objective function as : F(x) = t f \u2211 t=0 \u2211 i \u03932 i(t) (13) Where \u0393i(t) is the torque of joint i at the instant t: To validate the method proposed in section V, we will use a more complicated model : HOAP-3\u2019s legs. We consider only the lower limbs, so the upper parts of the body are equivalent to a mass on the chest (cf"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003360_gt2010-23802-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003360_gt2010-23802-Figure2-1.png",
+        "caption": "Figure 2. Test bearings: 4 and 5 pad tilting pad bearing configurations",
+        "texts": [
+            " The forced excitation into the system was delivered through the bearing housing using two orthogonally mounted electro-hydraulic shakers (exciters), which enabled the precise control of the excitation frequency and amplitude. The static operating position of the test bearing with respect to the test rotor was established with the static loader. As shown in Fig. 1 the static loader uses a hydraulic piston in combination with a pulley system in line with a soft spring pack. The electro-hydraulic exciters were used to super impose dynamic force on the static force (from the static loader), where the peak spectral amplitude for each frequency component was targeted at 2.5 microns. The test bearings are shown in Fig. 2 and consist of four different configurations of interest. Both 4 and 5 pad rocker pivot tilting pad bearing were tested where the 4 pad was tested in LBP configuration and the five pad bearing was tested in the LOP configuration. For each of these two bearing configurations, both 50% pad offset and 60% pad offset, were investigated with differing pad preload coefficients (shown in Fig. 2). Oil was delivered through the bearing housing oil-inlet port (Fig. 1) and then through oil nozzles positioned between tilting pads (Fig. 2). Several measurements were taken to both define an operating condition and also to calculate rotordynamic force coefficients. Operating conditions were established using the static load, oil inlet temperatures, rotor speed, and oil flow rate. Another measurement of interest was pad temperatures. For each bearing configuration three pad temperatures were measured near the trailing edge of the pad, offset from the edge by 12.25o. The calculation of rotordynamic coefficients required dynamic measurements (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure6-1.png",
+        "caption": "Fig. 6 The epicyclic gear train: (a) a section and (b) a three-dimensional view",
+        "texts": [
+            " Regarding the control of the whole robot, the moving platform twist t is specified, from which the angular velocities of the two motors are required. These are readily obtained from equations (43) upon inversion of matrix EJ, namely vin~E{1 J vout, E{1 J ~{r 1 {r6,5 1 r6,5 N \u00f045\u00de The foregoing relations were derived for the drive subsystem of limb I. Those of limb II are identical, except for the definition of vin, whose components are now q\u03073 and q\u03074. Figure 7 depicts a layout of one of the two EGTs. Each EGT is composed of one gear sun S, labelled 5 in Fig. 6(b); three planets Pi, i5 1, 2, 3, a typical one being labelled 6 in the same figure; one ring gear R, labelled 7 in that figure; one planet carrier C, labelled 8 in Fig. 6(b); and two reversing gears G, also included in the dynamic model. The RGB subsubsystem thus comprises two horizontal bevel gears for DS I and two for DS II. As indicated in Fig. 7, the axes of the planets lie a distance rC from AI. Given the simplicity of the geometry of the layout of the EGT, its contribution to the inertia matrix is derived from its kinetic energy T, rather than in the form of equation (38). For each unit T~ 1 2 I5v 2 5z 1 2 I7v 2 7z 1 2 I8v 2 8z3 1 2 I6v 2 6 z3 1 2 m6 rcv8\u00f0 \u00de2z2 1 2 IGv 2 6z2 1 2 mG rcv8\u00f0 \u00de2 \u00f046\u00de where I5, I6, I7, I8, and IG are the moments of inertia of the sun, the planets, the ring gear, the planet carrier, and each of the two identical reversing gears G respectively, all about their axes of symmetry, with Proc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003124_ecbs.2010.9-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003124_ecbs.2010.9-Figure2-1.png",
+        "caption": "Figure 2: (a) Free body diagram of a typical quadrotor [2]. (b) Alternative free body diagram of quadrotor helicopter [3].",
+        "texts": [
+            " The quadrotor helicopter is a highly nonlinear system, and the simulator used takes into account complex dynamical effects such as the moments of the vehicle, as well as the blade distortion due to high-rpm motion. As such, our controller uses this dynamical simulator as the plant, as its behavior is representative of the actual vehicle. Dynamical simulators provide additional behavioral factors over standard Eulerian laws of motion; notably, issues such as momentum are considered, which make a vehicle\u2019s simulated motion more realistic. By studying the free body diagrams of our quadrotor helicopter in (Fig. 2), it is possible to derive well-defined dynamics for the system. In [2] a proposed derivation of the nonlinear dynamics can be performed in North-East-Down (NED) inertial and body fixed coordinates [3]. As proofs are not realizable for the impending high level commands, simulations will be created based on the quadrotor helicopter dynamics described in [3] and [2]. Most importantly, the transfer functions relating the translational states as a function of the delta thrust generated by the rotors according to [2]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000469_14644193jmbd78-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000469_14644193jmbd78-Figure5-1.png",
+        "caption": "Fig. 5 Pressure distribution and bearing reactions",
+        "texts": [
+            " For 2R/ 2, engine bearings can be approximated to short-width bearings. An analytic solution can be obtained, where the pressure distribution is given as [31] p = 3U\u03b70\u03b5 Rjc ( 2 4 \u2212 \u03d22 ) sin \u03b1 (1 + \u03b5 cos \u03b1)3 (10) To determine the pressure distribution the following parameters are required. 1. The speed of entraining motion of the lubricant in the contact, u. This is given as U = 1/2\u03c9Rj. 2. The eccentricity ratio, \u03b5 = e/c, where e is the \u2018line of sight\u2019 excursion of the centre of the journal away from the fixed position of the bearing bushing (Fig. 5, and note that the film thickness, h = c(1 + \u03b5 cos \u03c4)). Note that c is the designed nominal radial clearance, typically 1/250 c/Rj 1/2000. Therefore, the unknowns required for the tribological analysis e and \u03c9 are obtained from the dynamic analysis, where e = Ri \u2212 Rj. A number of assumptions have been made, in addition to the short bearing assumption (2Rj/ > 2). First, the bearing bushing is considered to have a sufficient thickness, and made of a material of high elastic modulus in order to discount its deformation due to generated lubricant pressures",
+            " 221 Part K: J. Multi-body Dynamics JMBD78 \u00a9 IMechE 2007 at NATIONAL UNIV SINGAPORE on June 28, 2015pik.sagepub.comDownloaded from its value under atmospheric condition, \u03b70 (iso-viscous assumption). Finally, the solution provided here is isothermal, not taking into account the fall in the lubricant viscosity. With the pressure distribution determined, using equation (10), lubricant reactions (restoring forces) are usually given as load components Wx and Wz along and perpendicular to the line of centres (Fig. 5), respectively. Using the half-Sommerfeld\u2019s boundary conditions (pressure is generated in the region 0 \u03c4 \u03c0), then\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 Wx = U\u03b70 3 4c2 \u03c0\u03b5 (1 \u2212 \u03b52)3/2 Wz = U\u03b70 3 c2 \u03b52 (1 \u2212 \u03b52)2 (11) The resultant reaction is obtained as (Fig. 5) W = \u221a W 2 x + W 2 z The friction force in a lubricated conjunction is given by relative sliding motion of the two contacting surfaces. Here Uj refers to the surface velocity of the journal, where as Ub denotes the surface velocity of the bush bearing [31] F = \u222b 0 ( h 2 \u2202p \u2202x + \u03b70(Ub \u2212 Uj) h ) dx (12) Note that Ub = 0, and that the first term in the integral disappears with short bearing assumption as \u2202P/\u2202x = 0. Thus, noting that x = Rj\u03c4 , h = c(1 + \u03b5 cos \u03c4), \u03c8\u0307 = \u03c9 F = \u2212 \u222b\u03c0 0 Uj\u03b70 Rj c(1 + \u03b5 cos \u03b1) d\u03b1 = \u2212\u03c0\u03b70\u03c8\u0307R2 j c 1\u221a 1 \u2212 \u03b52 (13) The negative sign indicates opposition to the direction of motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000304_tac.2005.863493-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000304_tac.2005.863493-Figure2-1.png",
+        "caption": "Fig. 2. Cart-pole system.",
+        "texts": [
+            " This problem is a convex programming one and it can be reduced to a linear programming problem if one chooses the -norm for . Clearly the solution does not assure that the the condition is satisfied after the transient even for the initial condition . However, the condition can be assured for any if we consider a characteristic polynomial subject to the constraints (35). In this section, we report some experimental results obtained by applying a relatively optimal controller to the cart-pole system shown in Fig. 2. The system has one input (the current given to the electrical motor, proportional to the force applied to the cart) and two outputs (the position of the cart and the angle of the pole , measured by means of two encoders). The system has the following state vector: . A scheme of the system is reported in Fig. 3. The continuous-time model, linearized around a stable equilibrium point is where is the acceleration of gravity, the length of the pole, and the friction coefficients for the pole and the cart, respectively, and a coefficient that takes into account the whole mass of the cart and the pole and the proportionality between the input and the force applied to the cart (note that the dynamics of the pole does not influence that of the cart)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000966_la802967p-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000966_la802967p-Figure3-1.png",
+        "caption": "Figure 3. Capillary rise of isotropic-nematic (IN) interface up a solid flat wall (W). The director field (dashed lines) is assumed to be homogeneous. It makes an angle \u03c6 with the normal qW to the wall and an angle with the normal qM to the IN interface. The tangent to the profile and the horizontal make an angle R and \u03d1 is the contact angle of the profile with the wall.",
+        "texts": [
+            " The fluid interface rises up the wall because the surface tensions between the IW and the NW interfaces are not equal. We consider a nematic fluid domain that is assumed to be large enough not to be influenced by other boundaries and to be denser than the isotropic phase. For lyotropic liquid crystals, this is usually the case on account of the larger particle concentration in the nematic phase.2 The location of the interface is indicated by y(x), where y is the coordinate along the wall relative to a reference plane and x is the coordinate perpendicular to it; see Figure 3. The wall in the z-direction, i.e., perpendicular to the plane of interest, is supposed to be infinitely long so a quasi two-dimensional analysis suffices. The director field in the nematic domain is presumed to be rigid, that is, uniform, and to make a fixed angle \u03c6 with respect to the normal to the solid wall. So, formally we presume the limit \u03bb/lc f \u221e to hold. Below, we relax this condition and discuss what happens when \u03bb/lc is large but finite. The (contact) angle that the fluid-fluid interface makes with the wall we denote as \u03d1",
+            " The surface tensions of the three interfaces we denote \u03c3IN, \u03c3NW, and \u03c3IW, of which \u03c3IN and \u03c3NW are assumed to be of the Rapini-Papoular type,3 so of the generic form \u03c3 ) \u03b3 + w sin2 , with \u03b3 the usual (bare) surface tension and w the anchoring strength; is the angle between the field and the normal to the interface; see Figure 1. We now have to distinguish the NW interface from that of the coexisting isotropic and nematic phases. Let ) \u03c6 denote the angle of the director field and the normal to the NW interface and ) that at the IN interface. These two angles are related to each other if we presume the tangent y\u2032(x) \u2261 dy/dx to the profile of the IN interface (the meniscus) to be a known function of the distance x from the wall. Simple geometry then gives ) \u03c0/2 + arctan y\u2032 - \u03c6; see Figure 3. Hence, the surface tensions obey the relations \u03c3IN ) \u03b3IN +wIN sin2 ) \u03b3IN +wIN (cos \u03c6+ y\u2032sin \u03c6)2 1+ y\u20322 \u03c3NW ) \u03b3NW +wNW sin2 \u03c6 \u03c3IW ) \u03b3IW (1) As already alluded to in the Introduction, the sign of the anchoring strength w indicates the preferred type of anchoring. If w > 0, the preferred anchoring is homeotropic, and if w< 0, it is planar. Nematic dispersions of rodlike colloids for entropic reasons prefer planar anchoring, while in those of disklike colloids, the favored type of anchoring is homeotropic"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001451_978-0-387-09643-8-Figure6.13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001451_978-0-387-09643-8-Figure6.13-1.png",
+        "caption": "Figure 6.13: The Two Points on the Line Representing the Wall (adapted from [26]).",
+        "texts": [
+            " The error associated with zi can be calculated as follows: \u039bzi = ( \u2202zi \u2202yi )2 \u039byi By calculating the derivative in the above equation we get: \u039bzi = (\u2212Y0 y2 i )2 \u039by = Y 2 0 y4 i \u039by which shows how \u039bzi depends on the value of yi. Second, the angle between the robot and the wall (\u03b1) is calculated with the function: \u03b1 = sin\u22121 ( zl \u2212 zr D0 ) where D0 is the known distance between the two physical points pl and pr. Therefore, \u039b\u03b1 = ( \u2202\u03b1 \u2202zl )2 \u039bzl + ( \u2202\u03b1 \u2202zr )2 \u039bzr = \u239b \u239c\u239c\u239d 1\u221a 1\u2212 ( zl\u2212zr D0 )2 \u239e \u239f\u239f\u23a0 2 \u039bzl + \u239b \u239c\u239c\u239d \u22121\u221a 1\u2212 ( zl\u2212zr D0 )2 \u239e \u239f\u239f\u23a0 2 \u039bzr After simplifying the last equation we get: \u039b\u03b1 = D2 0 D2 0 \u2212 (zl \u2212 zr)2 (\u039bzl + \u039bzr ) Finally, we calculate two points on the line representing the wall as shown in Figure 6.13. Take the first point p1 at (0, zc) and the second point p2 at one unit distance from p1 along the wall which gives the point (cos \u03b1, zc + sin \u03b1): x1 = 0, z1 = zc x2 = cos \u03b1, z2 = zc + sin \u03b1 From these equations, the error for the two points will be: \u039bx1 = 0, \u039bz1 = \u039bzc \u039bx2 = sin2\u03b1 \u039b\u03b1, \u039bz2 = \u039bzc + cos2\u03b1 \u039b\u03b1 Now, we can write the error of ILSS1 as: error(ILSS1) = {\u039bx1 , \u039bz1 , \u039bx2 , \u039bz2} Notice that we can write the error in terms of \u039by, Y0, D0, yl, yc, and yr. For example, let\u2019s assume that \u039by = 1mm2, Y0 = 500mm, D0 = 300mm, and yl = yc = yr = 10mm (\u03b1 is zero in this case), then the error will be: error(ILSS1) = {0, 25mm2, 0, 25mm2} Now we analyze ILSS2 in a similar manner"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003724_2011-01-2169-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003724_2011-01-2169-Figure1-1.png",
+        "caption": "Figure 1. The physical model of a tire",
+        "texts": [
+            " And after knowing the magnitude and direction of total friction force, the longitudinal and lateral forces can be obtained eventually. In this paper, a set of test data with strong anisotropy of tire stiffness, in which the longitudinal stiffness is about three times larger than lateral stiffness, is used to verify the model results, which shows that the modification factor has good effect to enhance the accuracy of tire model at cornering/braking combination. SAE Int. J. Commer. Veh. | Volume 4 | Issue 1 (October 2011)84 The physical model of a tire [2] is simplified as Fig.1, assuming that the carcass of the tire can merely be deformed along the directions of X and Y axes translationally, without any angular deflection. The relationship between the tire force and velocity in the contact patch, under combined slip condition, is shown in Fig.2. F denotes the total friction force in the contact patch, the longitudinal and lateral components of which are represented by Fx, Fy respectively. The direction of wheel traveling speed, which is shown as V in the figure, gives a slip angle \u03b1 with respect to the central wheel plane"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002575_j.triboint.2011.10.005-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002575_j.triboint.2011.10.005-Figure1-1.png",
+        "caption": "Fig. 1. Hypoid gear contact.",
+        "texts": [
+            " Gear failures like wear, scuffing and micropitting strongly depend on lubrication conditions between gear tooth flanks. Many formulae have been developed to predict the lubricant film thickness in EHL contacts under various operating conditions. However, these methods were developed for applications where the surface velocities of the sliding partners are orientated in the same direction, like e.g. in spur gears. For bevel gears with hypoid offset (hypoid gears) for example the surface velocities v1 and v2 of the contacting partners are not orientated in the same direction. Fig. 1 shows a contact line on a hypoid gear flank with the surface velocity vectors v , 1 and v , 2 including the skew angle d. Due to the offset in a hypoid gear set the circumferential velocities of pinion and wheel are orientated in different directions. Therefore, the directions of the surface velocity vectors also differ. This results in additional components of sliding and sum velocity in lengthwise direction. Up to now it is unknown what influence the angle between the surface velocities and the additional components of sliding and sum velocity have on the formation of the lubricating film"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003968_cjme.2012.05.968-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003968_cjme.2012.05.968-Figure1-1.png",
+        "caption": "Fig. 1. Detailed view of gear tooth",
+        "texts": [
+            " According to HERTZ contact theory, we can conclude the computation formula of contact stiffness (K): 1/2 1 2 1 2 1 2 2 1 2 4 , 3 , , ( )(1 ) K ER R RR R R E EE E E \u00b5        (5) where R is combined curvature radius of the tooth surface in the contact point, E represents the combined Young\u2019s modulus of the materials involved in the contact process. The equation describing the normal deflection of the contact point is given as 2 2 55 , 4 m K \u03bd\u03b4       (6) where m represents the combined mass of the cylinder objects and \u03bd is the contact kinematic velocity of the object. In ADAMS, the contact damping can be described as a STEP function max max( ) STEP( , 0, 0, , ) , C d C\u03b4 \u03b4 \u03b4 \u03b4 (7) when normal deflection (\u03b4) is zero, the corresponding damping coefficient (C) is zero; similarly, as \u03b4 gets to the maximum, C also reaches its peak[14\u201315]. Fig. 1 depicts detailed view of 2D gear tooth and 3D gear tooth during the process of gear mesh. In order to obtain the contact force, we need to determine the penetration value between the two adjacent gear teeth. According to the non-penetration constraint principle, when the penetration value comes to its maximum, gears involved in the contact no longer penetrate into each other and turn to move towards the opposite directions under the action of contact force. For the process of gear transmission, 3D gear mesh is accomplished through the line contact of tooth surface while 2D gear mesh is realized by two points on the tooth profile"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001104_09544062jmes817-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001104_09544062jmes817-Figure2-1.png",
+        "caption": "Fig. 2 The geometric model of ball\u2013race interaction",
+        "texts": [
+            " Then, the dynamic behaviours of the rotor system supported by an angular ball bearing is investigated under the effects of different parameters, and the dynamic characteristics of ball bearings are computed considering the vibration of the rotor system. The study will provide evidence for accurately analysing the characteristics of the ball bearing and reasonably choosing the bearing operation parameters. In order to investigate the dynamic characteristics of a rotor ball bearing system, the non-linear contact forces should be determined. Figure 1 illustrates the relationship between the load {Fx , Fy , Fz , My , Mz} and displacement of an angular ball bearing {X2, Y2, Z2, \u03b8y , \u03b8z}. The geometric model of a ball and races is shown in Fig. 2. O is the mass centre and geometric centre of the outer ring, O2 is the mass centre and geometric centre of the inner ring, Ob is the mass centre and geometric centre of the ball, which are prescribed, respectively, in the inertial frame (X , Y , Z), the inner ring fixed frame (x, y, z), and the ball fixed frame (x, y, z). From the geometric interactions of ball and races, the contact deformation between ball and races can be obtained as follows \u03b41i = \u2223\u2223r i bc1 \u2223\u2223 \u2212 ( f1 \u2212 0.5 ) Dw (1) \u03b42i = \u2223\u2223r i bc2 \u2223\u2223 \u2212 ( f2 \u2212 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000337_tac.2006.884995-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000337_tac.2006.884995-Figure4-1.png",
+        "caption": "Fig. 4. LPRS and oscillations analysis.",
+        "texts": [
+            ") was derived for the case of the linear part given by a transfer function Wl(s) [7], [8]. The formula has the form of infinite series, which is convenient if the plant is given by the transfer function. This formula can be used for linear parts containing a time delay J(!) = 1 k=1 ( 1)k+1Re Wl(k!) + j 1 k=1 ImWl[(2k 1)!] 2k 1 (8) where Wl(s) = C(Is A) 1B or considered a given function. With the plant model available, the LPRS can be computed at various frequencies and the LPRS plot can be drawn on the complex plane (an example of the LPRS is given in Fig. 4). What is important is that the LPRS is a characteristic of the relay feedback system and can be computed from the plant model. Once the LPRS is computed, the frequency of the symmetric periodic solution can be determined from the following equation: Im J( ) = b 4c (9) which corresponds to finding the point of intersection of the LPRS and the horizontal line that lies below the real axis at b=(4c) (Fig. 4), and the equivalent gain [7], [8] of the relay (the gain of the relay with respect to the averaged motions) can be determined as kn = @u0 @ 0 =0 = 1 2ReJ( ) (10) which corresponds to the distance between the intersection point and the imaginary axis. Both formulas: (9) and (10) directly follow from the definition (6). With the formulas of the LPRS available, input-output analysis of the relay feedback system (Fig. 3) can be done in the same manner as with the use of the describing function method [12] (however, involvement of the filtering hypothesis is not needed any longer)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002564_1.4002379-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002564_1.4002379-Figure2-1.png",
+        "caption": "Fig. 2 MoS2 coated thrust ball bearings",
+        "texts": [
+            " Application to Fault Diagnosis of MoS2 Coated hrust Ball Bearing The fault diagnosis system for wear detection of the MoS2 oated thrust ball bearings is schematically shown in the Fig. 1. ibration signals measured from accelerometers are used to calulate the time-frequency representation of the time domain sigal. A feature from SPWV representation of the signal is calcuated to assess the condition of the coating. Simultaneous to easuring vibration, the coefficient of friction is measured as a unction of time. The information from both of these parallel meaurements is used to assess MoS2 coating performance and its egradation as a function of time. 3.1 Experimental Setup. Figure 2 shows the thrust ball bearng and its raceways. The bearings are made of 52,100 hardened teel raceways and each bearing include a steel cage and twelve .76-mm-diameter steel balls. The dimensions and coating specications of the coating are summarized in Table 1. Physical vapor eposition PVD technique is utilized to coat the races of the hrust ball bearings with MoS2. In a PVD process, the vaporized orm of the solid is deposited on the surface in an atomistic level. he deposition procedure is carried out in a vacuum or a very low ressure environment 43 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001477_j.1365-2109.2007.01895.x-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001477_j.1365-2109.2007.01895.x-Figure1-1.png",
+        "caption": "Figure 1 Schematic diagram of an electrolytic reactor.",
+        "texts": [
+            " Shrimp aquaculture wastewater characteristics Raw aquaculture wastewater was collected from the shrimp farm \u2018Ladang Ternakan Udang Harimau, JW Proprties\u2019, located at Kampung Assam Jawa, Kuala Selangor, Malaysia. The characteristics of the shrimp aquaculture wastewater are presented inTable1.The aquaculture wastewater was preserved for a maximum period of 2 weeks at a temperature of o4 1C, but above the freezing point in order to curtail biodegradation (APHA1985). The electrolytic reactor had a circular shape with the following dimensions: 300mm inner diameter and 450mmheightwitha liquid volume of 24 L as shown in Fig.1. A graphite rod 270mm in length and 60mm in diameter was used as an anode. Perforated stainless-steel sheets 270mm long, 50mm wide and having a thickness of 0.8mmwere used as cathodes.Two sets of anodes and cathodes were used during the electrolysis. In each set of electrode, two cathodes sheets were placed on either side of the graphite anode.The distance between the anode and the cathode was 20mm. A recti\u00a2er having an input of 230 V and a variable output of 0^20 V with a maximum current of 100Awas used as a direct current source"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002012_sled.2010.5542804-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002012_sled.2010.5542804-Figure1-1.png",
+        "caption": "Fig. 1. Supplied voltage deviation \u03b4vc and current response \u03b4ic in the qd-reference frame fixed to the rotor.",
+        "texts": [
+            " The voltage deviations \u03b4vcq, \u03b4vcd from the steady-state voltage vq,0, vd,0 are denoted by \u03b4vcq and \u03b4vcd respectively. It is then shown that the resulting current variations \u03b4icq, \u03b4icd after a time \u03c4 << min(\u03c4q, \u03c4d), with \u03c4q, \u03c4d the synchronous time constants of q- and d-axis, can be approximated by \u03b4icq = \u03c4 Lq \u03b4vcq and \u03b4icd = \u03c4 Ld \u03b4vcd, (1) independent of the stator resistor Rs as well as rotor speed \u2126. To estimate the rotor position \u03b8r, an auxiliary angle is used: the angle \u03b3\u03b4 of the voltage vector \u03b4vc(\u03b4vcq, \u03b4v c d) with respect to the q-axis, Fig. 1. As the set value of the voltage test vector is computed by the current controller, its direction is known in the stationary \u03b1\u03b2-reference frame and given by \u03b3\u2032 \u03b4. From this and by estimating the auxiliary angle \u03b3\u03b4 , an estimation of the rotor angle \u03b8r is obtained as \u03b8\u0303r = \u03b3\u2032 \u03b4 \u2212 \u03b3\u0303\u03b4 where x\u0303 denotes the estimation of x. This means that hereafter the dependence of the current response on \u03b3\u03b4 instead of \u03b8r should be studied from which an estimation of \u03b8r follows. The angle \u03b3\u03b4 can be introduced as follows"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001093_s106879980903012x-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001093_s106879980903012x-Figure1-1.png",
+        "caption": "Fig. 1. Fig. 2.",
+        "texts": [
+            " These seals suffer from severe drops of pressure and temperature resulting in increased force and thermal strains of sealing rings. A significant factor that influences the article operation is that turbomachines are to function in different operating modes. The seals must provide the reliable engine operation in all its operating modes. When designing these seals, it is necessary to take into account all factors listed. As an object of our investigation, use was made of the seal the scheme of which is presented in Fig. 1. On the sealing end face of the rotating ring, a number of complex-shaped microgrooves are made. At present, three types of microgrooves are used (Fig. 1, images A\u2013C). As the microgrooves are made in the circumferential direction, there exist alternate zones of increased and reduced pressure in the vicinity of their location; the \u201cnegative\u201d portion of the load-carrying capacity is equal in this case to the \u201cpositive\u201d one. The pressure value in the zone of reduced pressure is taken to be equal to that of saturated vapor, i.e., the lubricant coat fracture is observed. In such a way, the zones of increased pressure and zones of lubricant coat fracture are formed on the sealing ring; as a result, the \u201cnegative\u201d portion of the loadcarrying capacity becomes less than the \u201cpositive\u201d one, and the seal acquires the hydrodynamic loadcarrying capacity",
+            " In the analysis of processes because of cumbersome initial differential equations, a number of conventional assumptions dictated by some features of FV engine seal operation are taken, namely, a fluid is incompressible, a working fluid flow in the slot is laminar, and time variation of the clearance size is much less than its rated value. In order to calculate the characteristics of the hydrodynamic face seal with microgrooves, we developed a mathematical model based upon the finite volume method [1]. The essence of this method is BELOUSOV et al. RUSSIAN AERONAUTICS Vol. 52 No. 3 2009 336 that the entire volume of the annular face clearance is partitioned into sectors; each sector is divided, in its turn, into nine reference volumes. An example of such a volume isolated is presented in its cubical form in Fig. 2. Fig. 1. Scheme of the face seal with microgrooves: (1) fixed ring; (2) rotating ring; (3) secondary seals. Fig. 2. Example of sector portion division. Proceeding from the condition that the flowrates through the reference volume in the radial and circumferential directions are equal, we can find the pressure values at each point. If the centrifugal component of inertia forces is taken into account, the equation for determining the value of the pressure p at a point will have the form [2]: ( )1, 1, , 1 , 1 ,ij ij i j ij i j ij i j ij i j ij ij ijp B p C p D p E p F G A \u2212 + \u2212 + = + + + + + where the coefficients are the following 3 3 3 3 1/ 2, 1/ 2, , 1/ 2 , 1/ 2 ; 12 12 12 12ij i j i j i j i j h r h r h r h r A r r r r \u2212 + \u2212 + \u239b \u239e \u239b \u239e \u239b \u239e \u239b \u239e\u0394 \u0394 \u0394\u03d5 \u0394\u03d5 = \u2212 \u2212 \u2212 \u2212\u239c \u239f \u239c \u239f \u239c \u239f \u239c \u239f\u03bc \u0394\u03d5 \u03bc \u0394\u03d5 \u03bc \u0394 \u03bc \u0394\u239d \u23a0 \u239d \u23a0 \u239d \u23a0 \u239d \u23a0 3 1/ 2, ; 12ij i j h r B r \u2212 \u239b \u239e \u0394 = \u239c \u239f \u03bc \u0394\u03d5\u239d \u23a0 3 1/ 2, ; 12ij i j h r C r + \u239b \u239e \u0394 = \u239c \u239f \u03bc \u0394\u03d5\u239d \u23a0 3 , 1/ 2 ; 12ij i j h r D r \u2212 \u239b \u239e \u0394\u03d5 = \u239c \u239f \u03bc \u0394\u239d \u23a0 3 , 1/ 2 ; 12ij i j h r E r + \u239b \u239e \u0394\u03d5 = \u239c \u239f\u03bc \u0394\u239d \u23a0 ( )1/ 2, 1/ 2, 2;ij ij i j i jF r r h h \u2212 + = \u03c9\u0394 \u2212 3 2 3 2 2 2 , 1/ 2 , 1/ 2 , 1/ 2 , 1/ 2 3 3 ",
+            " Since the solution is found by the iteration method, we then insert the values of viscosity and density at each point of the computational grid at the corresponding values of pressure and temperature. Under the action of the force and thermal loads, the clearance acquires cone shape. Its waviness may appear due to nonuniformity of the pressure and temperature distribution in the circumferential direction. The calculations were conducted for the second type of microgrooves (eight ones are arranged uniformly along the circumference) with the variable depth in the range from 5.75 \u00b5m (at the periphery) up to 0.5 \u00b5m (on the inner groove radius) (Fig. 1, view B). The sealing parameters are the following: pressure difference is \u0394p = 5 MPa, rotor speed is \u03c9 = 1000 rad/s, working fluid is water. As a result of these calculations, we obtained the pressure distribution in the clearance (Fig. 3). The zones of increased pressure and lubricant coat fracture are clearly seen in this figure. If we optimize the geometry of sealing rings 1 and 2 (Fig. 1) and arrangement of secondary seals 3, the bending moments acting on the rings are counterbalanced. Owing to this fact, it is possible to provide the friction pair planarity; in other words, the clearance has no cone shape. BELOUSOV et al. RUSSIAN AERONAUTICS Vol. 52 No. 3 2009 338 The next step is the determination of thermal deformations by using the finite element ANSYS software. The combined heat-strength problem is solved in several stages: first, we find the temperature distribution; and then, the deformations of the friction pair rings are calculated on the basis of the heat problem solution"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001179_isit.2007.4557655-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001179_isit.2007.4557655-Figure4-1.png",
+        "caption": "Fig. 4. Illustration of moving distances and target set",
+        "texts": [
+            " A state s C S is represented as s = (s\u00b0, s1, s2), where s\u00b0, sl and x2 are defined in the same way as for the grid case, except that x, and ys now specify the horizontal and vertical coordinates instead of the indices. Similar to the grid case, we will show that by the time t = 6k + 1, for any state s C S, Pr{st = s} > ci1w8 for some positive constant cl. Consider a movement of the random walk that keeps the direction. Denote the distance traveled in the direction of movement, and orthogonal to the direction of movement at time t respectively by at and 13t, as shown in Fig. 4. Since nodes are randomly and uniformly distributed, it can be calculated that E(cat) = 4r I-t, and E(13t) 0. Let the turning probability p = ,ut = E)(r), and k For simplicity of exposition, we assume that the random walk starts from some horizontal state so with sl = ioLLa and s2 = jo,u. We can write out the expected location E(st) of the random walk at time t by assuming that at = /Io, and 73t = 0 for all t in which it moves to a state of a neighboring node. Then E(st) depends only on the times of turns and turning directions, and evolves according to the same nonreversible chain P on the grid as in Section 3",
+            " , k-1} and j C {0,1 , 2k-1}, we have Pr{E(s = s E(s') = i,ut, E(s Ja} > 24k Consider a target set of states S consisting of states of the same type in any given square of side ,ut in the network. Note that as n -) oc, the conditional distribution of s 2 given E(s 2) = j_t, is simply a shifted version of that given E(s 2) = 0. Therefore, by the property of modulo-2 operation and the Bayesian rule, it can be shown that at t = 6k, Pr{st C S} > 4k2 Now, consider for example an east state of node i, denoted by s. The west neighboring region of i contains a square of side 1-t, and let the set of east states in this square be the target set S, as depicted in Fig. 4. By the uniform convergence in the law of large numbers [8], when r = Q ogn, d' < n2r for any node i and direction 1 w.h.p., thus we have Pr{s6,+l= s} > Pr{s86k = } > 1 /2 2-5A C2 2 dmax nwTr2224k2 4n' It can be shown that the stationary distribution of P1 is approximately uniform, i.e., for any s C S, c3 < -Fs < C44n - 4n for some positive constants C3 and C4. Therefore, Pr{s6k+1 s} > C28 C17F- C4 C. Proof of Theorem 4.2 For any c > 0, from the proof of Lemma 4.1, there exists some T A Tmix(PI, C ()= (r-1 log(t/c)), such that for any t > T, we have wv(t) > c2 for all j, and 4n y t E4in )w() j (t) j Il wj(t) 4n 4n < ~1E jIy( 4n-Fjx-I1 + E7 1,vj(t) -4n-Fj x 1 (C26 lY(f) 1 + ' 4nx) = (II(O) IIj"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000519_s11071-006-9176-z-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000519_s11071-006-9176-z-Figure2-1.png",
+        "caption": "Fig. 2 Planar rigid pendulum with impact",
+        "texts": [
+            " (26) For the motion of the homogeneous disk on the plane of slope \u03b1, the following three cases are possible: Case 1. \u03b1 < \u03b10 (Equation (22)). The disk has no motion. Case 2. \u03b10 \u2264 \u03b1 < (Equation (26)). The disk has pure rolling (no sliding) motion. Case 3. \u03b1 \u2265 . The disk has rolling and sliding mo- tions simultaneously. 3.1 Simple pendulum and energetic coefficient of restitution The planar rigid pendulum with mass M and length L pivots around the frictionless pin joint O , and the tip impacts an inelastic horizontal surface S at point C (Fig. 2). The fixed cartesian reference frame x Oyz is chosen. The inclination of the pendulum with respect to the vertical axis Oy is the angle \u03b8 . At the impact point C , the coefficients of kinetic and static friction are \u03bck and \u03bcs , the coefficient of rolling friction is s, and the energetic coefficient of restitution is e\u2217. The ratio \u03c9s/\u03c9a of the separation angular speed \u03c9s = \u03c9(ts) and the approach angular speed \u03c9s = \u03c9(ta) at impact is calculated. The kinetic energy of the pendulum is T = 1 2 I\u03c9 \u00b7\u03c9, (27) Springer where I is the mass moment of inertia with respect to the joint O and \u03c9 = \u03c9k is the angular velocity of the pendulum",
+            " Using the kinematic coefficient of restitution e, the following equation can be written: e = \u2212 vCs \u00b7 vCa \u00b7 . (49) From Equations (47) and (49), one can compute the angular velocities of separation \u03c91s and \u03c92s for rods 1 and 2. The kinetic energy dissipated by friction T is T = T (ts) \u2212 T (ta), (50) Springer where T (ta) and T (ts) are the kinetic energies before and after the impact for the double pendulum. 4 Results 4.1 Simple pendulum In this section, results from computer simulations are presented. The rigid pendulum impacting a rough hor- izontal surface is shown in Fig. 2. The energetic coefficient of restitution is e\u2217 = 0.3. Figures 4\u20136 illustrate the ratio of the separation angular speed \u03c9s , the approach angular speed \u03c9a , and the coefficient e\u2217 as a function of the angle \u03b8 at the impact. The effect of the energy dissipated by friction at the impact between the pendulum and the horizontal surface S is shown. At small values of the angle \u03b8 , the contact sticks (\u03c9s = 0) if the coefficient of static friction is sufficiently large (\u03bcs \u2265 tan \u03b8 ) [3]. Also, for small angles \u03b8 , the work done by the friction force Ft is large in comparison with the work done by the normal contact force Fn "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001082_s11071-007-9205-6-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001082_s11071-007-9205-6-Figure1-1.png",
+        "caption": "Fig. 1 Six dependent spatial elastic coordinates",
+        "texts": [
+            " The elastic orientation of the cross-sectional frame is negligible and the equations will be finally expressed by the approximated strain field. In the motion equations, two additional elastic terms have appeared that would perish in the conventional nonlinear 3D Euler Bernoulli beam theory. In this theory, variation of strains and variation of elastic potential energy are derived from approximated strain field regarding negligible elastic orientation of beam cross-sectional frame [8]. 2 Flying support of the beam The flying support of the beam has six degrees of freedom. In Fig. 1, the frame of the flying support of the beam is denoted by FB . The inertial reference frame is shown by FI , the direction of whose third axis is in the negative direction of the gravity. Position, variation of position, velocity, and acceleration of the flying support are projected onto FI as expressions (1). FB and FI are orthogonal and right-handed coordinate reference frames and their axes are marked by 1, 2, and 3 in the figures to indicate, the first, second, and third axes, respectively",
+            " This provides us with motion equations much more accurate than that of the conventional nonlinear 3D Euler\u2013Bernoulli beam theory that gains variation of strains and variation of elastic potential energy from the approximated strains indicating negligible elastic orientation in the cross-sectional frame. Two additional elastic terms have appeared that would perish in the nonlinear 3D Euler Bernoulli beam theory. Hence, the motion equations of this paper are more accurate than that of the nonlinear 3D Euler\u2013 Bernoulli beam. The two new elastic terms in the motion equations create this accuracy. This enhances the nonlinear 3D Euler\u2013Bernoulli beam theory that derives variation of strains and consequently variation of elastic potential energy from the approximated strain field. In Fig. 1, FS is the cross-sectional frame after elastic deformation which is a curvilinear orthogonal righthanded coordinate frame whose first axis is tangent to the curve created by cross-sectional area centers. Its axes are marked by 1, 2, and 3 in the figures to indicate, the first, second, and the third axes, respectively. Crosssectional frame before elastic deformation is shown by FS0 . Center of cross-sectional area before and after elastic deformation is shown by S0 and S, respectively. Center of cross-sectional area before elastic deformation, S0, is located in distance s from B. The spatial independent variable s denotes the distance of S from B before deformation. It is a Lagrangian and not an Eulerian coordinate. Figure 1 simply expresses spatial elastic deformation of FS by six dependent coordinates of which only four are independent. Elastic displacement vector of S from B projected onto FB is shown by d, and elastic rotation transformation matrix projecting a vector from FB onto FS is shown by RSB in expressions (5) d = \u23a1\u23a2\u23a3u + s v w \u23a4\u23a5\u23a6 , R SB = \u23a1\u23a2\u23a31 0 0 0 cos \u03b3 sin \u03b3 0 \u2212 sin \u03b3 cos \u03b3 \u23a4\u23a5\u23a6 \u23a1\u23a2\u23a3cos \u03b2 0 \u2212 sin \u03b2 0 1 0 sin \u03b2 0 cos \u03b2 \u23a4\u23a5\u23a6 \u00d7 \u23a1\u23a2\u23a3 cos \u03b1 sin \u03b1 0 \u2212 sin \u03b1 cos \u03b1 0 0 0 1 \u23a4\u23a5\u23a6 . (5) Elastic deformation at S along the first axis of FB due to extension or compression is shown by u"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000918_s11465-008-0013-6-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000918_s11465-008-0013-6-Figure2-1.png",
+        "caption": "Fig. 2 One situation of subproblem S2-RRT",
+        "texts": [
+            "1 Solve subproblems using geometric method Subproblem S2-RRT is expressed as exp bj1h1 exp bj2h2 exp bj3h3 p~q, where j15 (u1,v1) and j25 (u2,v2) are zero pitch, unit magnitude revolute twists, and j35 (u3,0) is an infinite pitch, unit magnitude twist. p,q[R3 are two points. This subproblem corresponds to translate p along j3 by h3, and then rotate about the axis of j2 by h2, finally rotate about the axis of j1 by h1, thus, the final location of p coincides with point q. The problem solves joint variables h1, h2, and h3. Consider the situation of axes j1 and j2 intersecting as shown in Fig. 2. Define the vectors w5 (t2 r), u5 (p2 r), and v5 (t2 p). In triangle Drpt, cos a~ u3. {u\u00f0 \u00de u3k k uk k , wk k~ t{rk k~ q{rk k, and uk k are all known quantities, and h3~+ vk k can be solved by the law of cosines, wk k2~ uk k2zh23{2 uk kh3 cos a. Thus, point t can be solved by t5 r +w. Refer to [10], h1 and h2 can be solved through subproblem S2-RR. 3.2.2 Solve subproblems using algebraic method In subproblem S2-RRT, when axes are in general position, such as axes of j1 and j2, which are not coplanar, an algebraic method is easier than a geometric method"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002990_iros.2010.5649131-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002990_iros.2010.5649131-Figure2-1.png",
+        "caption": "Fig. 2. Impedance control with respect to the inertial coordinate",
+        "texts": [
+            "(4) with respect to the inertial frame are achieved by controlling the joint velocities \u03c6\u0307 as follows [9]: ffi\u0307 = J\u2217\u22121 jZ t 0 1 M i (Fh \u2212 Di\u0394x\u0307h \u2212 Ki\u0394xh) + x\u0308hddt \u2212 x\u0307gh ff , (5) where J\u2217 \u2208 R6\u00d7n is the generalized Jacobian matrix [10], x\u0307gh \u2208 R6 is the velocity of the gravity center of the entire system projected on the velocity of the end effector. In general, the contact in the target capture is a threedimensional phenomenon. However, It can be assumed as a one-dimensional phenomenon in the local viewpoint. In this case, the target mass is the GIE at the contact point [3]. Hereinafter, a one-dimensional satellite capture model is considered for simplicity. The dynamics model of a chaser robot under the impedance control is shown in Fig.2. In the figure, mi, di, and ki are mass, viscosity, and stiffness characteristics given by the control, respectively. From another viewpoint, the contact model combines the dynamics of each object in the contact state. The entire dynamics model during the contact can be represented as a coupled-vibration system as shown in Fig.3. In this figure, xh and xt are the chaser hand and target position, respectively. In this case, the penetration \u03b4 in Eq.(3) is equal to xh \u2212 xt. The equations of motion are as follows: mix\u0308h + (di + dc)x\u0307h + (ki + kc)xh \u2212 dcx\u0307t \u2212 kcxt = 0 mtx\u0308t + dcx\u0307t + kcxt \u2212 dcx\u0307h \u2212 kcxh = 0 } "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure1-1.png",
+        "caption": "Figure 1. Arrangement of the chosen coordinate system.",
+        "texts": [
+            " It is found that this number depends upon the trajectories of these counterweights. Various paths are considered for these counterweights, and practical applications to real mechanisms are discussed. Iy THIS article we shall examine the questions of determining the minimum number of counterweights necessary for full static, dynamic, and statico-dynamic balancing of plane linkages, and of deriving new means for balancing such mechanisms. Let us consider any linkage whose links move in one plane and attach to its base H the coordinate system Z O X (Fig. 1). Substitute for the motion of all inertia forces the motion of the resultant vector C'. and the resultant moment L., taking the center of mass of the moving links U. as the point of reduction. The mechanism is statically balanced if C~=O, (1) and dynamically balanced in the case for which L ~ = O . (2) When both conditions are satisfied simultaneously the mechanism is balanced staticodynamically. The balancing sequence cannot be fixed arbitrarily. The static balancing should either precede the dynamic or be carried out simultaneously with it"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002014_cefc.2010.5481434-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002014_cefc.2010.5481434-Figure1-1.png",
+        "caption": "Fig. 1. The proposed FT-FSPM motor topologies.",
+        "texts": [
+            " Over the years, some fault-tolerant (FT) machines, namely the switched reluctance [1] and permanent-magnet (PM) brushless [2] types, have been developed. However, they suffer from relatively low power density and weak mechanical structure, respectively. Recently, a new class of brushless motors with PMs located in the stator, termed as flux-switching PM (FSPM) motor, has been proposed [3]. The purpose of this paper is to propose a new modular FSPM (M-FSPM) motor, which can not only retain the advantages of the FSPM motor, but also solve its problems. A new 3-phase M-FSPM machine is proposed, as shown in Fig. 1. The key is the introduction of FT teeth between the adjacent stator poles. Thus, the individual windings are essentially isolated among phases, leading to significantly enhance the FT capability. Fig. 2 shows the no-load magnetic field distributions. It depicts that the M-FSPM motor can offer the nature of phase decoupling, which is essential for FT operation. By using finite element method, the proposed M-FSPM motor has been quantitatively compared with the conventional FSPM motor. An experimental M-FSPM motor is designed and prototyped for verification, as shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001275_j.compositesa.2009.02.014-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001275_j.compositesa.2009.02.014-Figure1-1.png",
+        "caption": "Fig. 1. Schematic diagram of heat/laser drawing apparatus.",
+        "texts": [
+            "0 wt% of TLCPs were spun by using the melt spinning machine with four holes and diameter of 1 mm and the throughput rate was controlled to adjust final denier of spun fibers. The speeds of feed and take-up rollers were 420 and 2550 m/min, respectively, and draw ratio was approximately 6.0. Various drawing methods, e.g., laser, heat, and heat/laser drawing processes, were applied to the spun fibers additionally in order to investigate drawing effects on mechanical and thermal properties of the TLCP/PEN composite fibers. Schematic diagram of the heat/CO2 laser drawing apparatus used in this study is shown in Fig. 1 and drawing conditions are summarized in Table 1. The draw ratio was calculated by the ratio of the speed of feed roller to that of take-up roller, and the diameter of both feed and take-up rollers was 10 cm. The wavelength and power of CO2 laser system (J48-2 Model, Synard Inc., USA) were 10.6 lm and 37.5 W, respectively. The beam diameter and the beam divergence angle of the CO2 laser were 5.0 mm and 0.1 mrad, respectively. Real generating power of the CO2 laser that irradiated the TLCP/PEN fiber was controlled by the laser power meter (POWER WIZARDTM 250 Model, Synrad, Inc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001806_analsci.26.417-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001806_analsci.26.417-Figure5-1.png",
+        "caption": "Fig. 5 Schematic structure of microfluidic polymer chips fabricated for evaluating the mixing efficiency between a sample solution and a carrier solution. a) A green solution, b) a red solution.",
+        "texts": [
+            " The selectivity coefficient values (log kPot NO3 \u2013X\u2013) exhibited in Table 3 indicate that the selectivity of the NO3 \u2013-ISE is in agreement with the Hofmeister series, which is correlated with the hydration energies of the anions. Effect\u202f of\u202f the\u202f structure\u202f of\u202f a\u202f microfluidic\u202f polymer\u202f chip\u202f on\u202f the\u202f mixing\u202f efficiency\u202f between\u202f a\u202f sample\u202f solution\u202f and\u202f a\u202f carrier\u202f solution The performance of a microfluidic polymer chip depends on the mixing efficiency between a sample solution and a carrier solution. Therefore, the effect of the structure of a microfluidic polymer chip and the flow rate of the sample and the carrier solutions on the mixing efficiency between a sample solution and a carrier solution was examined. Figure 5 shows a schematic diagram of three fabricated microfluidic polymer chips. A green dye solution (solution 1) and a red dye solution (solution 2) were flowed to channel a and channel b in the microfluidic polymer chip with a syringe pump, respectively. If solution 1 and solution 2 are mixed completely, the color formed by mixing solution 1 with solution 2 is dark green at an outlet tube connected with a detector chip. Table 4 shows the effect of the structure of a microfluidic polymer chip on the mixing efficiency between solution 1 and solution 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure7-1.png",
+        "caption": "Fig. 7 Schematics of the epicyclic gear trains",
+        "texts": [
+            "comDownloaded from v6~ 1 2r6,5 {v5z 1z2r6,5\u00f0 \u00dev7\u00bd \u00f041a\u00de v8~ 1 2 1zr6,5\u00f0 \u00de v5z 1z2r6,5\u00f0 \u00dev7\u00bd \u00f041b\u00de Substituting equations (39) into equations (41a) and (41b), while noticing that vp5v8 and vt5v6\u2013v8, leads to vp:v8~{ 1 2r 1zr6,5\u00f0 \u00de _q1z 1z2r6,5\u00f0 \u00de _q2 h i \u00f042a\u00de vt:v6{v8~ 1z2r6,5 2rr6,5 1zr6,5\u00f0 \u00de _q1{ _q2 \u00f042b\u00de Let vin~ _q1 _q2 \" # , vout~ vp vt which are related by the Jacobian EJ, J5 I, II, in the form vout~EJvin, J~I, II \u00f043\u00de with EJ given by EJ~ 1 Dr6,5 {r6,5 {Nr6,5 N {N , J~I, II \u00f044a\u00de where N and D are defined as N:1z2r6,5, D:2r 1zr6,5\u00f0 \u00de \u00f044b\u00de It should be noticed that, due to the symmetrical layout of the two EGTs with respect to the vertical plane equidistant from the two drive axes, the two Jacobians EI and EII are equal. Regarding the control of the whole robot, the moving platform twist t is specified, from which the angular velocities of the two motors are required. These are readily obtained from equations (43) upon inversion of matrix EJ, namely vin~E{1 J vout, E{1 J ~{r 1 {r6,5 1 r6,5 N \u00f045\u00de The foregoing relations were derived for the drive subsystem of limb I. Those of limb II are identical, except for the definition of vin, whose components are now q\u03073 and q\u03074. Figure 7 depicts a layout of one of the two EGTs. Each EGT is composed of one gear sun S, labelled 5 in Fig. 6(b); three planets Pi, i5 1, 2, 3, a typical one being labelled 6 in the same figure; one ring gear R, labelled 7 in that figure; one planet carrier C, labelled 8 in Fig. 6(b); and two reversing gears G, also included in the dynamic model. The RGB subsubsystem thus comprises two horizontal bevel gears for DS I and two for DS II. As indicated in Fig. 7, the axes of the planets lie a distance rC from AI. Given the simplicity of the geometry of the layout of the EGT, its contribution to the inertia matrix is derived from its kinetic energy T, rather than in the form of equation (38). For each unit T~ 1 2 I5v 2 5z 1 2 I7v 2 7z 1 2 I8v 2 8z3 1 2 I6v 2 6 z3 1 2 m6 rcv8\u00f0 \u00de2z2 1 2 IGv 2 6z2 1 2 mG rcv8\u00f0 \u00de2 \u00f046\u00de where I5, I6, I7, I8, and IG are the moments of inertia of the sun, the planets, the ring gear, the planet carrier, and each of the two identical reversing gears G respectively, all about their axes of symmetry, with Proc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003097_tasc.2010.2041208-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003097_tasc.2010.2041208-Figure1-1.png",
+        "caption": "Fig. 1. Schematic diagram of the giant magnetostrictive acceleration sensor.",
+        "texts": [
+            "2041208 magneto-mechanical strongly coupled FE model to calculate the input-output static characteristics of the magnetostrictive acceleration sensor and developed a giant magnetostrictive acceleration sensor [4]. In this paper, based on the energy functional, the dynamic model of the acceleration sensors built with rare earth giant magnetostrictive alloy has been founded, which includes eddy-current effects, and the time characteristics of the induced voltage for the sensor has been computed. In order to examine the validity of the proposed model, characteristics of the giant magnetostrictive acceleration sensor has been measured, too. Fig. 1 shows the schematic diagram of a giant magnetostrictive acceleration sensor system. It consists of a permanent magnet, iron yokes, an inductance coil and a giant magnetostrictive rod. The permanent magnet is used to supply the bias magnetic field. When a dynamic acceleration exerted on the magnetostrictive rod, according to the Newton\u2019s second law, mechanical force exerted on the giant magnetostrictive rod is known. And the magnetic flux density in it will change with the applied acceleration, according to the inverse magnetostrictive effect of the magnetostrictive materials",
+            " , , and are the displacement components of the rod in , , and directions, respectively. The elastic energy is as follows , where and are the stress and strain tensor, respectively. The relation between stress and strain tensor is , is the elastic module matrix. The work of the external forces is given by , where and represent the external surface force density and the external volume force density, respectively, the boundary of mechanical domain , and , the displacement complex vector. The electromagnetic field boundary in Fig. 1 should be at infinity. In practice, is assumed a finite boundary, if the solution of the electromagnetic field is consistent with one in a bigger domain, can be considered as the computed boundary of the electromagnetic field, shown in Fig. 3. If displacement current is ignored, the Maxwell relations can be written as (1) where is the magnetic field intensity, Js the source current density, here is none, Js the eddy current density, the electric field intensity, the magnetic flux density. The constitutive relation of the magnetic medium is (2) where is the permeability"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002507_iet-cta.2010.0589-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002507_iet-cta.2010.0589-Figure7-1.png",
+        "caption": "Fig. 7 Logarithmic quantiser (shown for positive input values only)",
+        "texts": [
+            " For each subsystem, it consists of four parts: the sensor terminal, the decoder terminal, the zero-order holder and IET Control Theory Appl., 2012, Vol. 6, Iss. 9, pp. 1304\u20131312 doi: 10.1049/iet-cta.2010.0589 a Response of the inner signal u to a square input b Response of the output signal y to a square input the continuous time plant. The continuous time analogue output of the plants are sampled every 0.15 s and quantised by logarithmic quantisers. The quantisation effect, as illustrated in Fig. 7, satisfies the sector bounded static non-linear property. In this example, they are bounded by [0.999, 1.001]. The size of the data packet is 10 bits. Then, the encoded signals are sent to the network. Meanwhile, the data packets at receiver sides are kept by zero-order holders. In Fig. 6, the last node is the external signals\u2019 source that generates unit impulse signals. The first five nodes are all stable LTI subsystems, and they are randomly chosen as follows G1 : 2 s3 + 6s2 + 11s + 6 ; G2 : \u22127 s3 + 7s2 + 14s + 8 G3 : 2 s + 10 ; G4 : 2 s2 + 6s + 5 ; G5 : \u22124 s2 + 3s + 2 Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-83-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-83-1.png",
+        "caption": "Figure 3-83: Section through modified main bearing.",
+        "texts": [
+            " It would also be appropriate to consult with oil and bearing suppliers. Cavitation erosion is an impact fatigue attack caused by the formation and collapse of vapor bubbles in the oil film. It occurs under conditions of rapid pressure changes during the crank cycle in internal combustion engines. The harder the bearing material the greater its resistance to cavitation erosion. Machinery Component Failure Analysis 159 Impact cavitation erosion occurring downstream of the joint face of an ungrooved whitemetal-lined bearing is shown in Figure 3-82. Figure 3-83 is a section through the main bearing. It illustrates the mechanism of cavitation and modifications made to the groove to limit or reduce damage. In Figure 3-84, observe the result of discharge cavitation erosion in an unloaded half-bearing caused by rapid movement of the journal in its clearance space during the crank cycle. Also of interest is Figure 3-85, a set of diesel-engine main bearings showing cavitation erosion of the soft overlay while the harder tin-aluminum is unattacked. The journal bearings shown in Figures 3-82 through 3-85 may be put back into service if cavitation attack is not very severe or extensive"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001877_1.3084237-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001877_1.3084237-Figure1-1.png",
+        "caption": "Fig. 1 Schematics of the line contact",
+        "texts": [
+            " The DC-CC-FFT algorithm is a hybrid fast Fourier transform based algorithm, which combines the discrete convolution\u2013FFT and the continuous convolution\u2013FFT methods. The proposed algorithms are used to solve three-dimensional displacement, contact pressure, and stresses for line contact problems. The results are compared with the other available algorithms from literature. The accuracy and efficiency of different algorithms are discussed. DOI: 10.1115/1.3084237 Keywords: elastic contact, FFT based algorithms, textured surface Introduction Contrary to ball bearings with point or elliptical contacts, spur ears and cam-roller follower systems Fig. 1 are engineering xamples of line contacts. The cylinder-on-flat contact pairs and lock-on-ring contact pairs are laboratory testing examples of line ontacts, which are frequently used in the reciprocating testing rig 1 and in the block-on-ring testing machines 2 , respectively. epending on applications, a cylinder is often chosen instead of a all specified in the ASTM G133 standard. The surface treatents, such as coating and textured surfaces, are introduced to nhance performances, such as gear endurance life and cam-roller cuffing resistance, whereas the roughness and/or slight damages re inevitable reality for contact surfaces that influence component ife",
+            " These algorithms thus can transform a 3D contact solver, which was originally developed for point contacts or nominally flat contact problems 34 studied in Ref. 35 by a statistical approach , to deal with the SPLC problems. The proposed methods have been validated with APRIL 2009, Vol. 131 / 021408-109 by ASME of Use: http://asme.org/terms a w o 2 i t I i t a c h c p h f e r n r i u u t w m o v i T w c F c w p 0 Downloaded Fr Hertzian contact. The contact problems of a textured surface ith and without coating have been solved with proposed methds as demonstrations. Theory In this paper, half-spaces Fig. 1 are assumed for both bodies n contact where the contact width is small enough in comparison o the radius. Therefore, the superposition principle is applicable. n the contact length direction, the crown shape is not considered n this study. A coating can be considered in the model, although he discussion in this section uses uncoated half-spaces as an exmple. Nonuniform friction coefficients can be specified over the ontacting surface for the consideration of different regimes of ydrodynamic lubrication, boundary lubrication, or solid-to-solid ontact"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001396_j.ijengsci.2008.08.002-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001396_j.ijengsci.2008.08.002-Figure1-1.png",
+        "caption": "Fig. 1. Annular axisymmetric stagnation flow on a moving cylinder. (a) The inner cylinder rotates with angular velocity X and move axially with velocity W. The outer cylinder is fixed with fluid injected towards the inner cylinder. (b) Cross section showing streamlines in the annular region.",
+        "texts": [
+            " In the present paper, the stagnation flow in the annular region between two cylinders is studied. Fluid is injected inward radially from a fixed outer cylinder towards an axially translating and rotating inner cylinder. Such finite geometry is more realistic for the convective cooling of a moving rod [6]. The problem also models externally pressure-lubricated journal bearings which are quite attractive for high speed and miniature rotating systems [7\u201311]. The aim of the present paper is to find a similarity stagnation flow for the annular region. Fig. 1 shows a vertical inner cylinder (shaft) of radius R rotating with angular velocity X and moving with velocity W in the axial z-direction. The inner cylinder is enclosed by an outer cylinder (bushing) of radius bR. Fluid is injected radially with velocity U from the outer cylinder towards the inner cylinder. Assuming end effects can be ignored (long cylinders), the flow is axisymmetric about the z-axis. The constant property continuity equation and the constant property Navier\u2013Stokes equations in axisymmetric cylindrical coordinates are: "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000528_s11768-006-6048-5-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000528_s11768-006-6048-5-Figure2-1.png",
+        "caption": "Fig. 2 Rotational actuator to control a translational oscillator. Fig. 3 and 4 shows the state trajectories when the system is at rest and experiencing a disturbance",
+        "texts": [
+            " Reducing the mesh size, we have \u3008w\u0307T L(t)\u03c3L(x),\u03c3L(x)\u3009\u03a9 = lim \u2016\u0394x\u2016\u21920 (ATA) \u00b7 w\u0307L(t) \u00b7 \u0394x, \u3008w\u0307T L(t)\u03c3L(x)f(x),\u03c3L(x)\u3009\u03a9 = lim \u2016\u0394x\u2016\u21920 (ATB)\u00b7wL(t)\u00b7\u0394x, \u30082 uL 0 \u03d5\u2212T(v)Rdv, \u03c3L(x)\u3009\u03a9 = lim \u2016\u0394x\u2016\u21920 ATC \u00b7 \u0394x, \u3008wT L(t)\u2207\u03c3L(x)g(x)\u03d5( 1 2 R\u22121gT(x)\u03c3T L(x)wL(t)),\u03c3L(x)\u3009\u03a9 = lim \u2016\u0394x\u2016\u21920 (ATD) \u00b7 wL(t) \u00b7 \u0394x, \u3008wT L(t) 1 4\u03b32 \u2207\u03c3L(x)k(x)kT(x)\u2207\u03c3T L(x)wL(t),\u03c3L(x)\u3009\u03a9 = lim \u2016\u0394x\u2016\u21920 ATwT L(t)E \u00b7 wL(t) \u00b7 \u0394x, \u3008hT(x)h(x),\u03c3L(x)\u3009\u03a9 = lim \u2016\u0394x\u2016\u21920 (AT \u00b7 F )\u0394x, This implies that (27) can be converted to w\u0307L(t) =\u2212(ATA)\u22121wL(t)ATB \u2212 (ATA)\u22121ATC +(ATA)\u22121ATDwL(t)\u2212(ATA)\u22121ATwT L(t) \u00b7EwL(t) \u2212 (ATA)\u22121ATF. (35) This is a nonlinear ODE that can easily be integrated backwards using final condition wL(tf ) to find the least-squares optimal NN weights. In this example, we will show the power of our NN control technique for finding nearly optimal finite-horizon H\u221e state feedback controller for the Rotational/Translational Actuator shown in Fig. 2. This was defined as benchmark problem in [45]. x\u0307 = f(x) + g(x)u(t) + k(x)d(t), |u(t)| 2, zTz = x2 1 + 0.1x2 2 + 0.1x2 3 + 0.1x2 4 + \u2016u(t)\u20162 q, \u03b5 = me/ \u221a (I + me2)(M + m) = 0.2, \u03b3 = 10, f = [ x2 \u2212x1+\u03b5x2 4 sin x3 1 \u2212 \u03b52 cos2 x3 x4 \u03b5cos x3(x1\u2212\u03b5x2 4 sin x3) 1 \u2212 \u03b52 cos2 x3 ]T , g = [ 0 \u2212\u03b5 cos x3 1 \u2212 \u03b52 cos2 x3 0 1 1 \u2212 \u03b52 cos2 x3 ]T , k = [ 0 1 1 \u2212 \u03b52 cos2 x3 0 \u2212\u03b5 cos x3 1 \u2212 \u03b52 cos2 x3 ]T . Here the state x1 and x2 are the normalized distance r and velocity of the cart r\u0307, x3 = \u03b8, x4 = \u03b8\u0307. Define performance index V (x(t0), t0) = \u03c6(x(tf ), tf ) + tf t0 (hT(x)h(x) +2 u 0 \u03d5T(v)dv \u2212 \u03b32\u2016d(t)\u20162)dt"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000809_robot.2008.4543556-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000809_robot.2008.4543556-Figure5-1.png",
+        "caption": "Fig. 5. Kinematic model of four-wheeled vehicle",
+        "texts": [
+            " Controlling the angular velocity of each driving axle also compensates the longitudinal slip of wheel. To discuss the nonholonomic kinematic model, the following assumptions are considered: 1) the distance between wheels (generally called wheelbase and tread) are strictly fixed; 2) the steering axle of each wheel is perpendicular to the terrain surface; 3) the vehicle does not consist of any flexible parts. 1) Kinematic model with slip angle: A kinematic model of a four-wheeled vehicle including the slip angle of vehicle and lateral slips of wheel is shown in Fig. 5. In this model, each wheel has a steering angle \u03b4i and slip angle of wheel \u03b2i. The slip angle of wheel, which measures a wheel lateral slip, is calculated by same equation as (1) using the longitudinal and lateral linear velocities of wheel, vix and viy , as follows: \u03b2i = tan\u22121(viy/vix) (4) The subscript i denotes the wheel ID as shown in Fig. 5. The position and orientation of the center of gravity of the vehicle is defined as (x0, y0, \u03b80), while (xi, yi) gives the position of each wheel. vi is the linear velocity of each wheel.. l is the longitudinal distance from the center of gravity of the vehicle to the front or rear wheel and d defines the lateral distance from the center of gravity of the vehicle to the left or right wheel. Here, based on previously defined assumptions, l and d are constant values. 2) Nonholonomic constraints: The nonholonomic constraints are expressed by the following equation, taking into account the lateral slip: x\u03070 sin \u03c60 \u2212 y\u03070 cos \u03c60 = 0 (5) x\u0307i sin \u03c6i \u2212 y\u0307i cos \u03c6i = 0 (6) where, \u03c60 = \u03b80 + \u03b20, and \u03c6i = \u03b80 + \u03b4i + \u03b2i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003837_wst.2013.509-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003837_wst.2013.509-Figure1-1.png",
+        "caption": "Figure 1 | Schematic diagram of the M20 membrane filtration system. 1. Feed tank, 2. Piston pump (rotary lobe pump \u2013SRU), 3. Pressure pump, 4. Pressure-inlet gauge, 5. Membrane module, 6. Pressure-outlet gauge, 7. Pressure control",
+        "texts": [
+            " The supernatant was collected after 20 min of sedimentation and filtered through Whatman No. 1001-185 filter paper (pore size of 11 \u03bcm) on a Buchner funnel with slight suction as the prefiltration step. The filtrate was used for UF or NF filtration experiments. Membrane filtration experiment The filtration experiments were carried out in a plate-andframe membrane module (DSS LabStak M20, Alfa Laval, Nakskov, Denmark), which is a cross-flow membrane filtration system. The schematic diagram of this system is shown in Figure 1. The membrane sheets were stacked and compressed together with spacer plates in a vertical frame in the membrane module. The system was operated under recirculating flow. The flow rate of the feed solution was 480 L/h during the experiments. The diameter of each membrane sample was 20 cm, and the effective area during the filtration was 0.0174 m2 for each piece of membrane. Four pieces of each membrane were used in the experiments. The influences of different operating pressures were investigated in this work"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003883_j.proeng.2011.08.646-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003883_j.proeng.2011.08.646-Figure1-1.png",
+        "caption": "Figure 1 Structure of FS",
+        "texts": [
+            " For the FS multi-objective optimization with unclear preference, it is found that the results gained by general optimization methods are still unsatisfactory in preference and decisionmaking. Therefore, it is necessary to present an effective method to solve the above problems for FS. For the FS multi-objective optimization, the leakage (Q) and operation life (expressed by using wear ratio V generally) are taken as the two conflicting objectives. By combining fuzzy theory and the game theory, the multi-objective optimization results of different preference requirement for FS could be gained. The structure of FS is shown in Fig 1. The FS is to process a set of finger beams in the thin slice and make the finger slices staggered close together to cover the adjacent interstice. The multiple finger slices and two cover plates are assembled with the rivet tightly. The seal is fitted over the rotating shaft or rotor with a small amount of clearance or interference, depending on the application. The fluid through the seal is impeded by the staggered fingers/pads as well as the radial contact between the rotor and the FS feet. From the FS geometric characteristics, we find that the angle of finger stems (\u03c6), thickness of the finger element (s), finger length (The finger length is control by the radius of base circle (r) if the shape-curve of finger stems is involute), the height of finger foot (h) and the angle between finger stems (\u03b4) are the main parameters for FS"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001219_s12541-009-0102-4-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001219_s12541-009-0102-4-Figure9-1.png",
+        "caption": "Fig. 9 Load application points and displacement measurement points for estimating the thrust and radial compliances of the C-axis feed system",
+        "texts": [
+            " Further, the axial compliance of the X-axis, Y-axis, and Z-axis feed systems was obtained from an axial load applied to the front of each table and the average value of axial displacements measured at the four measurement points. As the measurement results, the normal, lateral, and axial compliances of the X-axis, Y-axis, and Z-axis feed systems were found to be 0.0390 \u00b5m/N, 0.3020 \u00b5m/N and 0.0006 \u00b5m/N; 0.0013 \u00b5m/N, 0.0061 \u00b5m/N and 0.0010 \u00b5m/N; and 0.0479 \u00b5m/N, 0.0520 \u00b5m/N and 0.0076 \u00b5m/N, respectively. Fig. 9 shows two load application points and four displacement measurement points used to estimate the thrust and radial compliances of the C-axis feed system. The four displacement measurement points are located at the circumference of the C-axis table. The thrust compliance of the C-axis feed system was estimated from a normal load applied to the center of the C-axis table and the average value of thrust displacements measured at the four measurement points, and the radial compliance of the C-axis feed system was evaluated from a radial load applied to the side of the C-axis table and the average value of radial displacements measured at the four measurement points"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001343_ichr.2007.4813886-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001343_ichr.2007.4813886-Figure2-1.png",
+        "caption": "Fig. 2. Representation of zmp",
+        "texts": [
+            " The dynamic model equation allows to compute the joint torques \u0393(t) knowing the joint angle, velocity and acceleration : \u0393(t) = NE(q(t), q\u0307(t), q\u0308(t), t) (11) In [5], the Newton-Euler method is used for computing the dynamic model of a 3-D robot thanks to a two-recursionsalgorithm. This paper deals with optimization applied to 2- D robot, therefore we adapt the Newton-Euler method for computing the joint torques of a 2-D robots. For humanoid robots, the computation of the Zero Moment point (ZMP) give information about the balance. [6] defines ZMP as the point zmp, on the contact surface, where the moment is equal to zero Mzmp = 0 (cf. fig 2). If this point stays in the base of support, the robot maintains its equilibrium. The zmp location depends on the joint angle q(t), velocity q\u0307(t) and acceleration q\u0308(t). The double pendulum (cf. fig.3) is used, in section IV, as a simple model. We define the initial value qi = [0,0] to get the foot position : (x = 0, y = 0). The final joints value is computed to do a step lenght of d : q f = [ Acos ( d 2L ) ,2\u00d7Acos ( d 2L )] (12) Let us consider one parameter per joint. So, in the case of Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.31-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.31-1.png",
+        "caption": "Fig. 1.31. On the stability of a trailer (Exercise 1.4).",
+        "texts": [
+            " Another effect of this reduction of the average cornering stiffness is that when the vehicle moves at a speed lower than the critical speed, the originally stable straight ahead motion may become unstable if through the action of an external disturbance (side wind gust) the slip angle of the caravan axle becomes too large (surpassing of the associated unstable limit-cycle). This is an unfortunate, possibly dangerous situation! We refer to Troger and Zeman (1984) for further details. Exercise 1.4. Stability of a trailer Consider the trailer of Fig. 1.31 that is towed by a heavy steadily moving vehicle at a forward speed V along a straight line. The trailer is connected to the vehicle by means of a hinge. The attachment point shows a lateral flexibility that is represented by the lateral spring with stiffness Cy. Furthermore, a yaw torsional spring and damper are provided with coefficients c~, and k~,. Derive the equations of motion of this system with generalised coordinates y and ~o. Assume small displacements so that the equations can be kept linear"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003749_s1068366612020110-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003749_s1068366612020110-Figure5-1.png",
+        "caption": "Fig. 5. Schematic of calibration of contact load control system of SI 03M test machine.",
+        "texts": [
+            " (4) The error of the slippage coefficient control sys tem is determined by the program variation of its value and the corresponding variation of the rotational fre quencies of the specimen and counterspecimen. The error is determined at the points 0, 25, 50, and 75% at a preset rotational frequency of the specimen drive of 3000 rpm; it should be no more than 2% of the mea sured value. (5) The error of the contact load control system is determined using DOSM 3 1U and DOSM 3 2U master proving rings. Figure 5 shows the schematic for this error check. Before the check, the contact loading system is calibrated. The load is assigned in the manual control mode. The fine adjustment of the load is car ried out using the rod (see Fig. 5). The measurement error of the contact load FN is \u22482%. (6) The error of the system of measuring the bend ing load is determined using the DOSM 3 1U master proving ring. The check involves the following two modes: \u2014when the bending load is directed \u201cUPWARDS\u201d; \u2014when the bending load is directed \u201cDOWNWARDS\u201d. The arrangement of the devices and proving rings is shown in Fig. 6. Before calibration, the bending load control system is calibrated. The load is assigned in the manual control mode. The fine adjustment of the load is carried out using the rod (see Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001558_lars.2008.26-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001558_lars.2008.26-Figure2-1.png",
+        "caption": "Fig. 2. Helicopter main axis and attitude angles.",
+        "texts": [
+            " In order to develop a model that can be accurate enough, rigid body dynamics are coupled with simplified rotor dynamics, as presented in this section. The rigid body equations for a six-degree of freedom helicopter are given by the Newton-Euler equations, adapted from [2]: u\u0307 =\u2212wq+ vr\u2212gsin(\u03b8)+ Fx m , (1) v\u0307 = wp\u2212ur+gsin(\u03c6)cos(\u03b8)+ Fy m , (2) w\u0307 = uq\u2212 vp+gcos(\u03c6)cos(\u03b8)+ Fz m , (3) p\u0307 = 1 Ixx (\u2212qr(Iyy\u2212 Izz)+Mx), (4) q\u0307 = 1 Iyy (\u2212pr(Izz\u2212 Ixx)+My), (5) r\u0307 = 1 Izz (\u2212pq(Ixx\u2212 Iyy)+Mz), (6) In the equations above, u = x\u0307, v = y\u0307 and w = z\u0307 are the linear velocities and p, q and r are the angular velocities about the x, y and z axis as shown in Fig. 2. The forces and moments acting on the i\u2212th axis are Fi and Mi, respectively. The acceleration due to gravity is denoted by g, and m is the helicopter mass. Ixx, Iyy and Izz are the moments of inertia around the x, y and z axis. For reduced-scale helicopters, the cross products of inertia can be neglected [2]. The forces and moments Fi and Mi are generated by the main rotor, the tail rotor and by the aerodynamic effects in the fuselage, the vertical fin and the horizontal stabilizer. In order to improve the rigid-body model, simplified rotor dynamics can be coupled to it"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001455_6.2008-6462-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001455_6.2008-6462-Figure4-1.png",
+        "caption": "Figure 4. Turn anticipation geometry",
+        "texts": [
+            " DL = RPP sin (\u03c7PP \u2212 \u03c7AW ) (4) DV = RPP cos (\u03c7PP \u2212 \u03c7AW ) cos \u03b3PP sin (\u03b3PP \u2212 \u03b3AW ) (5) 4 of 14 American Institute of Aeronautics and Astronautics Therefore, right and up deviations from the active leg are considered positive. In this paper we consider fly-by waypoints, i.e. points which define navigations legs and require an anticipated turn to smooth the transition to the next leg. At a certain distance from AW, denominated distance of turn anticipation (DTA), the reference path changes to a circular trajectory tangent to the active and the future leg defined between the active waypoint and future waypoint (FW). Figure 4 illustrates in the local horizontal plane the geometry of the desired transition. The radius of turn (RT ) is calculated based on ground speed Vg and desired turn rate (\u03c7\u0307T ) according to RT = Vg |\u03c7\u0307T | (6) The desired turn rate is positive for turns to the right. The difference of desired tracking between two legs, \u2206\u03c7T , shown in Figure 4, is calculated as \u2206\u03c7T = \u03c7FW \u2212 \u03c7AW (7) DTA is obtained from DTA = RT tan \u2223\u2223\u2223\u2223\u2206\u03c7T2 \u2223\u2223\u2223\u2223 (8) Thus, the circular transition path to be followed when the distance to AW is less than DTA is given by6 DLT = sgn (\u2206\u03c7T )RT (1\u2212 cos (\u03c7\u2212 \u03c7FW )) (9) \u03c7 = \u03c7AW + \u03c7\u0307T t (10) The operation given by Eq. (10) initiates with t = 0 when the horizontal distance to AW is less than DTA and terminates when \u03c7 = \u03c7FW . The waypoint sequence is also updated when the distance to AW is less than DTA: the active waypoint changes to past waypoint and the future waypoint changes to active waypoint"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002158_j.advengsoft.2009.12.019-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002158_j.advengsoft.2009.12.019-Figure6-1.png",
+        "caption": "Fig. 6. CS disk parameterization.",
+        "texts": [
+            " (H, I, J) Bore, inner web, and outer web width The widths of the disk bore, inner web, and outer web are all input as a ratio of the disk rim width. The T-Axi Disk inputs PT1, PT2, and PT3 set these dimensions. (K,L) Inner rim definition The inner rim thickness distribution is defined by a fourth order line segment fit through three control points of the form: t\u00f0r\u00de \u00bc C1r4 \u00fe C2r3 \u00fe C3r2 \u00fe C4r \u00fe C5: \u00f019\u00de The five coefficients C1\u2013C5 define the general shape of the curve and can be found from the prescribed slope of the line at the two end points and the known value of t at each control point. As shown in Fig. 6, two of the control points at the disk bore and the inner web have already been defined. The third control point is assumed to be located at the radial midpoint between these two points. The thickness of the disk at the middle control point is specified in a non-dimensionalized form as T-Axi Disk input S12. Typically this input can have values between 0.0 and 1.0, with 0.0 being the case where the thickness at the middle control point equals the thickness at the inner web and 1.0 being the case where the thickness at the middle control point equals the thickness at the disk bore"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002010_s0022-2836(66)80096-7-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002010_s0022-2836(66)80096-7-Figure1-1.png",
+        "caption": "FIG. 1. Comparison of standard and modified immunoelectrophoresis runs. Fast = faster- and slow = slower-moving allelic antigens. Antiserum is placed in the central trough.",
+        "texts": [
+            " It is generally difficult to obtain precise measurements of migration rates by immunoelectrophoresis, mainly because of the imprecision of measuring the distance from the origin to the line bisecting the arc formed at the new antigen source. This difficulty was compounded here by the small differences expected in distances travelled by allelically determined proteins. However, by modifying the methods described by Grabar & Williams (1955) and Scheidegger (1955), slight differences could be resolved repeatedly. Usually the two antigens to be compared were placed in single origins separated by a trough filled with antiserum. In our system, four wells were placed as illustrated in Fig. 1. After a run was completed, serum was added, and the new positions of the antigens were marked by the two pairs of arcs of antigen-antibody precipitates that developed. The critical measurement was that of the length of a line intersecting at right angles the two lines bisecting the arcs on each side of the trough. Ifboth antigens moved at the same rate, the length of this connecting line would be the same whether measured on the right or left side of the trough, i.e. no matter which antigen was placed in the more forward well, the difference in positions of the arcs would not change"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000700_13506501jet265-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000700_13506501jet265-Figure1-1.png",
+        "caption": "Fig. 1 Geometry of gear teeth contact",
+        "texts": [
+            " The results obtained are plotted to show the variation of pressure distribution, film shape, maximum pressure and minimum film thickness with the couple-stress parameter at various points along the line of action for various values of gear ratio. ENTRAINMENT VELOCITY It is well known that involute profile originates from a base circle such that the normal to the involute profile is tangential to the base circle. When two gears are in mesh, two involute profiles come in contact such that the common normal to the involute profiles is also the common tangent to the two base circles. Since the common tangent to the base circles is fixed, therefore, all the points of contact are located on it. Figure 1 shows the geometry of gear transmission. The centers of the driver pinion and driven gear are O1 and O2, respectively. Let a pair of teeth be in contact at a particular instant when another pair of teeth comes into engagement simultaneously. The initial contact of this new pair of teeth takes place when the flank of the pinion tooth touches the tip of the gear tooth. This occurs at point A where the addendum circle of the gear crosses the pressure line, i.e. the common tangent to the base circles of the gear and pinion",
+            " The point of contact O moves from A along the involute profiles of the pinion and gear teeth in mesh and its path follow the pressure line. At a particular point (say C), the contact breaks between the pair of teeth which were already in engagement prior to the pair under consideration and hence, only a single pair of teeth bears the total load. Again, as the contact point reaches the position D, another pair of teeth comes into engagement. Finally, the contact between this pair of teeth breaks at point B where the pinion addendum circle intersects the pressure line. The portion CD, not shown in Fig. 1, along the line of action (AB) represents the span in Proc. IMechE Vol. 221 Part J: J. Engineering Tribology JET265 \u00a9 IMechE 2007 at Freie Universitaet Berlin on May 12, 2015pij.sagepub.comDownloaded from which only a single pair of teeth remains in contact. Hence, the total load is shared by more than one pair of teeth in the portions AC and DB. The point P is the pitch point. The length of the line of action is AB = AP + PB AP = \u221a r2 a2 \u2212 r2 2 cos2 \u03c6 \u2212 r2 sin \u03c6 (1a) PB = \u221a r2 a1 \u2212 r2 1 cos2 \u03c6 \u2212 r1 sin \u03c6 (1b) where, r1, r2 are the pitch circle radii and ra1, ra2 are the addendum circle radii of the pinion and gear, respectively. In order to determine the distances AC and CD, it is required to find the contact ratio (CR) CR = AB (\u03c0m cos \u03c6) (2) where, m is the module. Hence AC = (CR \u2212 1)\u03c0m cos \u03c6 (3a) CD = (2 \u2212 CR)\u03c0m cos \u03c6 (3b) In the present analysis, several positions of the contact point O from A to B are considered and these positions are represented by the distance s measured with respect to the pitch point P along the line of action, as shown in Fig. 1. At each position s of O, transient EHL line contact analysis is performed with the rolling direction x being tangential to the involute profiles at the point of contact, and hence, perpendicular to AB. Along the line of action, the curvature of the two involute profiles keeps changing at the point of contact. The radii of curvature of the pinion and gear tooth profiles are given below as functions of the position, s, of contact point O with respect to P and the x-coordinate along the rolling direction relative to O [4] R1(x, s) = r1 sin \u03c6 + s \u2212 rb1 r1 sin \u03c6 + s (4) R2(x, s) = r2 sin \u03c6 \u2212 s + rb2 r2 sin \u03c6 \u2212 s (5) where, r1, r2 are the pitch circle radii and rb1, rb2 are the base circle radii of the pinion and gear, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003695_2011-01-2700-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003695_2011-01-2700-Figure1-1.png",
+        "caption": "Figure 1. Newcastle University 75 mm centre distance back-to-back gear test rig",
+        "texts": [
+            " The authors concluded that all three techniques (acoustic emission, vibration analysis and ODA) not only offer an earlier indication of gear damage, but can also predict the onset of failure which enhances the results concluded [11,12,18, and 19] for the spur gears transmission system investigated. An existing back-to-back gear test rig based at Newcastle University was used for this research. The rig comprises a 75 mm centre distance back-to-back test gearbox connected to a slave gearbox in a layout shown in Figure 1. The test gears, each consisting of a pinion and a wheel, are lubricated by a pressurised spray of Atlas Copco Roto Z lubrication oil, pumped from a 60 litre tank located beneath the rig bed. In order to ensure that the lubrication oil fed to the gearbox is free of particles, it is passed through a 125\u03bcm suction filter, a magnetic filter and two 10\u03bcm filters (Maxiflow MX. 1518.4.10) before reaching the gears. The used oil from the test gearbox is returned to the tank by gravity through a return pipe"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.3-1.png",
+        "caption": "Fig. 3.3. Two-link planar arm at a boundary singularity",
+        "texts": [
+            " Unlike the above, these singularities constitute a serious problem, as they can be encountered anywhere in the reachable workspace for a planned path in the operational space. Example 3.2 J = [ \u2212a1s1 \u2212 a2s12 \u2212a2s12 a1c1 + a2c12 a2c12 ] . (3.40) det(J) = a1a2s2. (3.41) For a1, a2 = 0, it is easy to find that the determinant in (3.41) vanishes whenever \u03d12 = 0 \u03d12 = \u03c0, \u03d11 being irrelevant for the determination of singular configurations. These occur when the arm tip is located either on the outer (\u03d12 = 0) or on the inner (\u03d12 = \u03c0) boundary of the reachable workspace. Figure 3.3 illustrates the arm posture for \u03d12 = 0. By analyzing the differential motion of the structure in such configuration, it can be observed that the two column vectors [\u2212(a1 + a2)s1 (a1 + a2)c1 ]T and [\u2212a2s1 a2c1 ]T of the Jacobian become parallel, and thus the Jacobian rank becomes one; this means that the tip velocity components are not independent (see point a) above). Computation of internal singularities via the Jacobian determinant may be tedious and of no easy solution for complex structures"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003888_s00021-012-0105-2-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003888_s00021-012-0105-2-Figure6-1.png",
+        "caption": "Fig. 6. Minimum point x0 < x1. The point (x0, u0) \u2192 (\u2212\u221e, 0) as \u03b31 increases to \u03b30 1 , when \u03b32 is fixed and the same for both curves",
+        "texts": [
+            "3) can be reduced to elliptic integrals and asymptotically estimated according to established theory. In these ways, the net force F = \u03c3\u03bau2 0 can be obtained to arbitrary accuracy. This does not of course obviate Theorem 4.1 as a general description of the qualitative behavior of solutions. If instead of (4.2) we have \u03b31 \u2264 \u03b32, we exploit the symmetry in the problem to interchange \u03b31 and \u03b32 in the above relations. Case 2. The minimum lies outside the interval between the plates. A positive minimum outside the interval between the plates appears if and only if \u03c0/2 < \u03b31 < \u03b30 1 . See Fig. 6. There then holds x0 < x1 = x2\u22122a, and the discussion becomes somewhat more intricate, as one sees by observing that for the data \u03b30 1 and \u03b30 2 = \u03b32, as achieved by the barrier I described in Sect. 3, the net force vanishes regardless of plate separation. (We note of course that in this example \u03b30 1 varies with the separation). Nevertheless we can obtain meaningful asymptotic estimates, using a similar procedure as above. For such a configuration, the minimizing abscissa x0 occurs at a distance a0 > 2a from x2, and we have \u221a \u03baa0 = \u03c82\u222b 0 cos \u03c4d\u03c4\u221a \u03bau2 0 + 2 (1 \u2212 cos \u03c4) "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001299_14399776.2009.10780988-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001299_14399776.2009.10780988-Figure11-1.png",
+        "caption": "Fig. 11: Friction force measurement principle",
+        "texts": [
+            " To measure the Stribeck-curve over a broad range also the high translational velocities, where pure hydrodynamic lubrication is present, is in the focus of the experimental studies. To achieve velocities up to 10 m/s, a cylinder tube test rig with high dynamic force compensation was developed, see Fig. 10. The cylinder tube is mounted inside the piston rod housing, and is driven by a crank mechanism. Due to the high resulting inertia forces (up to 30 kN at 2000 rpm) a mass force compensation gear box is designed. Fig. 10: High velocity friction force test rig for pneumatic seals Figure 11 shows the measurement principle. The cylinder tube is in motion while the friction force at the test seal which is placed at the fixed test piston is measured using a piezoelectric force sensor. This principle allows measuring the friction at the test seal separately without having an influence of other frictional forces that occur. To apply a pressure at one side of the seal a second piston is used. The advantage of this concept is the non-varying pressurized capacity that guarantees a constant system pressure"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002022_1077546310384002-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002022_1077546310384002-Figure2-1.png",
+        "caption": "Figure 2. The oil injection system of the active TPJB.",
+        "texts": [
+            " The unbalance force vector caused by the eccentricity of the disc, Q1, and the nonlinear oil film force vector at the bearings, Q2, can be described as Q1 \u00bc m0 2e cos t m0 2e sin t m0g 0 0 0 mAg 0 mBg T Q2 \u00bc 0 0 0 0 FA x FA y FB x FB y T : \u00f012\u00de A typical set of parameter values for a rotor-bearing system with active TPJB used in the numerical simulations are shown in Table 1. The actively lubricated bearing worked out by Santos (1993) is adopted here. It is built of four tilting pads, in a load-on-pad configuration, as shown in Figure 2(a). The control action over the rotating shaft is made by injecting oil into the bearing gap through five machined bores (Santos and Nicoletti, 2001) in the pads (Figure 2(b)). By coupling servo-valves to the pads in the vertical and horizontal directions (Figure 2(a)), the pressure of the injected oil can be controlled. Thus, the hydrodynamic pressure and temperature distribution in the gap (main mechanism of bearing load capacity) may be altered among the different pads. Rotating shaft vibration can also be attenuated with help of control techniques as can be seen in Santos and Scalabrin (2000). The active bearing is composed of four tilting pads: two in the horizontal and two in the vertical directions (x- and y-axis of the inertial reference frame, see Figure 2(a)). The oil film pressure distribution on each pad is obtained by solving the modified Reynolds\u2019 equation @ @x h3i @pi @x \u00fe @ @y h3i @pi @y 3 l0 Fi x, y\u00f0 \u00de pi \u00bc 6U @hi @x \u00fe 12 @hi @t 3 l0 Fi x, y\u00f0 \u00dePi \u00f013\u00de where hi is the gap between the rotor and the ith pad, is the oil dynamic viscosity, l0 is the orifice\u2019s length, U is at East Tennessee State University on June 5, 2015jvc.sagepub.comDownloaded from the linear velocity of the rotor surface, Pi is the injection pressure in the ith pad (active lubrication), which can be obtained by solving the Equation (9), Fi\u00f0x, y\u00de is a positioning function of the orifices on the pad surface, and Fi x, y\u00f0 \u00de \u00bc d 2 0 =4 2, for 2 d 2 0 =4 0, for 2 4 d 2 0 =4 \u00f014\u00de where \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x xi0\u00de 2 \u00fe \u00f0 y yi0\u00de 2 q , \u00f0xi0, yi0\u00de is the position of the orifice center on the ith pad surface",
+            " Thus Fx \u00bc P4 i\u00bc1 Fni cos \u2019i \u00fe i\u00f0 \u00de Fti sin \u2019i \u00fe i\u00f0 \u00de , Fy \u00bc P4 i\u00bc1 Fni sin \u2019i \u00fe i\u00f0 \u00de \u00fe Fti cos \u2019i \u00fe i\u00f0 \u00de , 8>>< >>: \u00f018\u00de where \u2019i is the angular position of the ith pad pivot around the bearing (see Figure 3(b)). The hydraulic system consists of a reservoir, two pumps and two servo-valves. One of the pumps supplies the conventional lubrication to the bearing. The other pump feeds the injection system with high pressurized oil. The purpose of the servo-valves is to control the pressure at which oil is injected, through the pad bores, into the bearing gap (see Figure 2(b)). The dynamics of the oil flow through servo-valves I and II can be described by \u20acQVI \u00fe 2 VI !VI _QVI \u00fe !2 VI QVI \u00bc !2 VI KVI uI t\u00f0 \u00de \u20acQVII \u00fe 2 VII !VII _QVII \u00fe !2 VII QVII \u00bc !2 VII KVII uII t\u00f0 \u00de ( \u00f019\u00de where QV is the oil flow through the servo-valve under null loading (open ducts), V is the damping factor of at East Tennessee State University on June 5, 2015jvc.sagepub.comDownloaded from the servo-valve, !V is the eigen-frequency of the servo-valve, KV is the gain of the servo-valve, u\u00f0t\u00de is the control signal, and indices I and II identify the servo-valves connected to the pads in the horizontal and vertical directions respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003004_s1052618812030041-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003004_s1052618812030041-Figure1-1.png",
+        "caption": "Fig. 1. Parallel mechanism with different kinematic routes.",
+        "texts": [
+            " The key novelty of this paper lies in determining kinematic and forcing screws acting on the output unit (platform) and corresponding to the maximal generalized speeds, maximal generalized forces, and max imal forces in passive pairs. Let us consider a parallel structure manipulator with six degrees of freedom and six kinematic routes. Four of them consist of two spherical joints and one linear drive. One transmission route consists of a rota tional drive, a rectilinear pair, and two Hooke\u2019s joints (universal joints). One more kinematic route con tains a rotary driving apparatus and five non drive rotating pairs. This structure was chosen to demonstrate different cases of placement of forcing and kinematic screws. Figure 1 shows that the screws acting on the platform from four kinematic routes R1, R2, R4, and R5 have a zero parameter. The parameter of the screw R3 is infinite. The sixth kinematic route is of particular interest since the parameter of the screw corresponding to this route is not equal to zero and infinity. To find the forcing screw R6 acting on the platform from the sixth kinematic route, it is necessary to consider a group of individual screws corresponding to this route. The screw R6 is reciprocal to the group of individual screws from the non drive (passive) kinematic pairs",
+            " Thus, we consider only the coordinates proportional to the coordinates of the screw R6 so that Note that the parameter p\u03a9 of the kinematic screw \u2126 is not equal to 1/pR6, where pR6 is the parameter of the screw R6. If any of the coordinates is equal to zero, e.g., r6x = 0, then \u03bdx = 0. If p\u03a9 < 1, it is assumed that \u03bd = \u03bdmax, where \u03bdmax is the maximal speed of the platform\u2019s rectilinear motion. If p\u03a9 > 1, it is assumed that \u03c9 = \u03c9max, where \u03c9max is the maximal rotary speed of the platform. Now, it is possible to find the gen eralized speed . If this speed is higher that the admissible value , the mechanism is close to the type 1 singularity. Let us analyze the numerical parameter corresponding to Fig. 1. Let the admissible generalized speed of the sixth kinematic route be = 1 s\u20131. The maximal rotary and linear speeds are \u03c9max = 1 s\u20131 and \u03bdmax = 1 m/s. The screw R6 at the point O' has the following Plucker coordinates (1, 1, 1/2, \u20131, 1, \u20133); the corresponding Plucker coordinates for \u2126 will be as follows: (0.301, \u20130.301, 0.904, \u20130.301, \u20130.301, \u23af0.15). The relative moments which do not depend on the point of application will be as follows: In this case, = 3.992 s\u20131 is larger than the admissible value (1 s\u20131)",
+            " If pF > 1, it is assumed that f = fmax, where fmax is the maximal external force acting on the platform. If pF < 1, it is assumed that m = mmax, where mmax is the maximal external moment acting on the platform. Now it is possible to find the reaction k6R6 of the kinematic route on the passive pairs. If the components k6r6 or of this reaction are larger than the admissible values r6adm and , the mechanism is close to the type 2 singularity. Let us consider the numerical example corresponding to Fig. 1. Assume that the admissible values are r6adm = 1 kN and = 1 kN m. The maximal external force is fmax = 1 kN, and the maximal external moment is mmax = 1 kN m. The Plucker coordinates of the kinematic screw \u21266 relative to the point O' are (1, 0, 1/2, \u20131/2, 0, 3/2). The corresponding Plucker coordinates of the forcing screw F are (0.147, \u2013 0.442, \u20130.884, \u20130.294, 0, \u20130.147). The relative moments (irrespective of the reference point) will be as follows: In this case, k6 = 1.669, k6r6 = 2.504 kN, = 5",
+            " In the positions close to the type 2 singularity, the coefficient k6 takes on larger values, but in the case of proximity to the type 1 singularity, the relative moment + + + + + is small, and the generalized force m6 can also be small. Thus, it is possible to formulate the criterion for proximity to the type 3 singularity. If the components k6r6 and of the reaction k6R6 in passive pairs of the kinematic route are larger than the admissible values r6adm and , and if the generalized force m6 is smaller than the admissible force m6adm, the mechanism is close to the type 3 singularity. Let us consider the numerical example corresponding to Fig. 1. Assume that the admissible values r6adm and , the maximal external force fmax, the maximal external moment mmax, the kinematic screw \u21266, the forcing screw F, the coefficient k6, and the components of the reaction k6r6 and are the same as those in the previous case. The mechanism is in proximity to the type 2 singularity. It is necessary to find out if the mechanism is also close to the type 3 singularity. Assume that the admissible generalized force is m6adm = 1 kN m. The relative moment will be + + + + + = \u20131, and the generalized force of m6 = 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000098_j.msea.2006.01.079-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000098_j.msea.2006.01.079-Figure6-1.png",
+        "caption": "Fig. 6. Designation of axis inbound orientation axis of rotation at edge centre of base unit panel pair for (a) 0\u25e6 inbound orientation and (b) 45\u25e6 inbound orientation.",
+        "texts": [
+            " Standard tenile test pieces were used, and testing occurred at a strain rate f 1000 mm/min. Five trials were carried out for each test direcion, with maximum test strains being approximately 0.6. Fig. 4 hows a 5th order polynomial expression fitted to elastic moduli ata. High levels of symmetry are exhibited with similar values ig. 3. Resulting trellis type of deformation when straining along the material ias. D.S. Price et al. / Materials Science and Engineering A 420 (2006) 100\u2013108 103 rotation at the centre of the edges of two base unit panel pairs as shown in Fig. 6. Fig. 6(a) shows an example of a 0\u25e6 inbound orientation with the yarn directions being coincident with the direction of travel. Fig. 6(b) shows a 45\u25e6 inbound orientation, with the yarn directions forming an angle of 45\u25e6 with the direction of travel. Each carcass was impacted from 0 to 360\u25e6 in 15\u25e6 increments. Each impact was carried out using a bespoke 2-wheel ball launch device, which worked by pushing the ball in between two counter-rotating rollers, which launched the ball onto a steel plate. Each impact was recorded using a Photron Fastcam \u2013 Ultima APX 120K HSV camera operating at 10,000 frames per second which was triggered to initiate the storage of images into a microprocessor linked to a laptop through a firewire connection",
+            " Throughout a typical impact the ball undergoes hoop strain at maximum deformation whereby a great circle can be drawn that is parallel to the plate and coincides with the maximum tangential deformation point. The great circle undergoes a diametric increase throughout ball deformation. The impact testing protocol was designed so that the majority of strain throughout the impact will be evident within the central 4 base unit panel pairs which have been used to construct the axis of rotation as detailed in Fig. 6. This idea can be combined with the mechanical properties of fabrics detailed in Section 3 to provide reasoning for the orientation dependent impact properties observed throughout experimentation. Fig. 11(a) gives details of the 0\u25e6 inbound orientation impact showing a single solid arrow which represents the yarn directions, with the dotted line showing the plane of hoop strain. The panels that are shown in view have their yarn directions orientated coincident with the plane of hoop strain. Throughout deformation, the hoop strain experienced by the ball is analogous to applying uniaxial tension along the dotted line"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000034_pedes.2006.344353-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000034_pedes.2006.344353-Figure1-1.png",
+        "caption": "Fig. 1. Test bench layout.",
+        "texts": [
+            " Chapter IV and V analyzed the operating performance of a three-phase squirrel-cage induction motor under unbalanced sinusoidal (0.96% unbalance, 2.9% Total Harmonic Distortion (THD)) and balanced non-sinusoidal (inverter) supply voltages (23.7% THD) through a real load test and the economic analysis included in the chapter VI. The testing of the 4 pole, 400V, 5 HP squirrel-cage induction motor covered stator and rotor copper losses, slip, power factor and iron losses were measured. The experimental set up for testing induction motor is shown in Fig. 1. This experimental set-up is capable to handle non-sinusoidal and unbalanced quantities, again with the aim of treating inverter supplied drives. The input power, input voltages and currents are measured using voltage and non-contacting current probes, and the signals are digitally processed by the power quality (PQ) analyzer (Fluke 434). A DC generator is used for applying variable load to the motor. (3) True Voltage Unbalance Factor (VUF), the positive and negative sequence components of a three-phase, sinusoidal and unbalanced voltage system can be calculated exactly without the application of the Fortescue transformation in the complex plane, which only the RMS line-to-line voltages are required [3]: V U F V V (3) where V+ and V represent the voltages of the positive and negative sequence components, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003904_s11249-012-0067-9-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003904_s11249-012-0067-9-Figure1-1.png",
+        "caption": "Fig. 1 Schematic diagram of the fiber wobbling method",
+        "texts": [
+            " We revealed the dependency of the properties on the interaction between the lubricant molecules and the disk surface [4\u20137]. For example, the effective viscosity measured with the lubricant confined in the nanometer-scale gap widths was enhanced by a few to dozens of times larger than those in the bulk state. In addition, the elasticity, which was negligibly small in the bulk state, increased drastically at gaps that were equivalent to a few molecular sizes. A schematic diagram of the FWM is shown in Fig. 1. The optical fiber probe used as the shearing probe is fabricated to have a spherical end shape and is set perpendicular to the sample surface to shear the lubricant film with its tip. The shear force acting on the probe tip is measured by detecting the deflection of the probe using the fiber as a cylindrical lens to focus a laser spot onto a position sensitive detector (PSD). A change in the position of the laser spot corresponds to deflection of the probe. The shear force S. Itoh (&) Y. Hamamoto K"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002565_gt2011-46492-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002565_gt2011-46492-Figure2-1.png",
+        "caption": "Figure 2. Initial brush seal domain: 3 rows, 2 mm bristles, 45\u00b0 lay angle, 142 \u03bcm wire diameter, 10 \u03bcm bristle spacing, 1 mm backing ring clearance",
+        "texts": [
+            " Even though these deflections change the interstitial gap sizes along the bristle length, and lead to inter-bristle and bristle-backing ring contact, an idealised pack configuration gives a reasonable indication of the expected flow and temperature fields within the pack. Further, forces induced by the cold build of the seal play a part in changes which take place in the pack geometry as soon as pressurisation commences. arrangement of an initial test domain of 3 rows (along bristle axis) An initial test domain was created that comprised a bristle pack of 3 rows ( n ) and an exit region under a backing ring, shown in Fig. 2. The bristles of diameter ( d ) 142\u03bcm, were inclined (\u03b8 ) at 45\u00b0 to the radial direction, and spaced 10 \u03bcm apart (\u03b4 ) all the way along their length (L), which was 2 mm. As such, bristle contact and vanishing fluid volumes were avoided. Bristle spacing was based on pack thickness measurements at the root for a range of test seals and is typical of real bristle packs with high packing efficiency. A length, L, of 2 mm would be entirely unrealistic in a brush seal, exhibiting high stiffness and no ability to accommodate shaft excursions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001430_6.2008-7157-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001430_6.2008-7157-Figure3-1.png",
+        "caption": "Figure 3. ELAL flight 1862: actuator fault/failure and structural damage (adapted from8,23)",
+        "texts": [
+            " The aircraft crashed 13km east of Schiphol airport into an apartment building in Bijlmermeer, a suburb of Amsterdam. Further details on the incident can be found in.23 Unknown to the flight crew, the inboard fuse-pine that held engine no. 3 to the pylon broke due to fatigue. This caused no. 3 engine and its pylon to separate from the right wing shortly after takeoff causing damage to the leading edge of the right wing. The shedding of engine no. 3 from the right wing in an outboard and rearward direction resulted in a collision with no. 4 engine (see Figure 3), causing it and its pylon to separate from the wing. Figure 3 illustrates the estimated damage to the right wing. (The amount of damage on the wing leading edge after the separation of pylon no. 2 from an accident at Anchorage on March 31, 1993 (Figure 4), is indicative of the damage probably inflicted on the El Al 1862 wing23). In summary, the damage and its effect on the ELAL flight 1862 aircraft after the separation of engines no. 3 and 4 are: eThe role of the fuse pin is to allow the engine to separate from the wing under a strong impact load that occurs in the event of a crash or hard landing in order to protect the fuselage from engine fire",
+            " right wing leading edge damage causing changes to the wing aerodynamic; 3. the right inboard aileron and remaining spoilers 10 and 11 are less effective due to the airflow disruption resulting from the damage to the leading edge of the right wing; 4. limited roll control due to the loss of the outboard ailerons. 5. a mass loss of about 10 tonnes due to the separation of two engines. 6. hydraulic and pneumatic system no. 3 and 4 pressure loss. 7. loss of effectiveness of control surfaces due to the loss of hydraulic systems no. 3 and 4 (see Figure 3). 8. lateral CG displacement due to the loss of the engines. 9. degraded lateral control due to lower rudder lag as a result of hydraulic pressure loss. 10. a (positive) yawing moment to the right due to asymmetric thrust from engines no. 1 and 2. Early FTC studies on ELAL flight 1862 by Maciejowski & Jones9 showed that it is possible to control the crippled aircraft (although in,9 an exact model of the damaged aircraft was assumed to be available). It is 4 of 18 American Institute of Aeronautics and Astronautics 5 of 18 American Institute of Aeronautics and Astronautics important to highlight that, for the duration of the incident, the flight crew was unaware that engines no"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001443_1.30762-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001443_1.30762-Figure1-1.png",
+        "caption": "Fig. 1 Elevation over azimuth gimbal axes; subscript p denotes platform axes (LoS denotes line of sight).",
+        "texts": [
+            " Given the particular threat of global terrorism to the airline industry, future DIRCM systems should be small, lightweight, and costeffective for inclusion on commercial aircraft, which invariably limits the number of available axes for sightline control. A minimum of two axes are necessary to attain full hyperhemispherical FOR. The outer gimbal (OG) axis is usually constrained to rotate with respect to the platform z axis (through azimuth angle ) and the inner gimbal (IG) with respect to the OG y axis (through elevation angle \"). This is shown in Fig. 1. When the outer and inner axes are orthogonal, the sightline can be positioned to any orientation in space, as the IG and OG axes form an ideal orthogonal basis. However, when the inner gimbal angle approaches the nadir (defined as \" 90 deg) (i.e., the sightline becomes aligned with the axis of rotation of the outer gimbal), the pointing system experiences the loss of a degree of freedom known as gimbal lock and, as a result, the sightline cannot be arbitrarily positioned [2]. To guarantee tracking close to the nadir, the outer gimbal has to follow a large (theoretically infinite) acceleration demand, beyond the capability of any practical servomechanism"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001447_robot.2007.363852-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001447_robot.2007.363852-Figure4-1.png",
+        "caption": "Fig. 4. Force Propagation",
+        "texts": [
+            " To analyze the force propagation when the user grabs the suture, we need to compute the forces acting at each point from the grabbed point to the start point and to the end point of the suture. We define different scenarios as follows: fi- i fi- fi- i fi+ i fi A. Condition A Assume the user grabs point Pi+ with one hand. If Imin < 1i < ln,,. There is no propagation of the user input force fl1 from point Pi+1 to Pi. All the user input force has been converted to the internal forces along the suture. B. Condition B Assume the user grabs point Pi+, with one hand. If the expected segment length 1' > or l' < 1min, we need to adjust the segment length to ImaX or Imin (see (a) of Fig. 4). and check if the BHV of the rope has any overlap with the line segments. When two suture segments are detected to be at a distance d < 2r from each other, then, an equal (but opposite) displacement vector is applied to each segment along. This displacement is just long enough to take the segments out of collision, with a slight \"safety margin\". Hence, each node is shifted away by r-d 2 + \u00a3/2. If a collision occurred, during real time simulation, we need to compute new velocities of mass points which are in'volved in the collision",
+            " fp and f10 can be obtain from the following equations: fp = (f7 ei)ei fni (fhuem)eCn. (16) (17) where e-,,1 =xf , ei can be obtain from equation (5). Using the same method as above, we can derive the user input force propagated at each point of the suture. C. Condition C In this condition, we assume the user is pulling two points,Pk and Pi, of the suture. The method is almost the same as in condition B. But we need to do the propagation computation twice, first starting from point Pi, and then starting from point Pk (see (b) of Fig. 4). A. Collision Detection and Management First, we build a bounding-volume hierarchy (BVH) from the bottom-up representing the shape of the rope at successive levels of detail. This method is similar to the method proposed in [1L6] To find the self-collisions of the rope we explore twx o copies of the BVH from the top down. Whenever two BVHs (one from each copy) are found to not overlap, we know that they cannot contain colliding segments, and hence, we do not explore their contents When two leaf spheres overlap, the distance between the two centers of the nodes is computed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003434_978-94-007-4902-3_36-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003434_978-94-007-4902-3_36-Figure1-1.png",
+        "caption": "Fig. 1 Bevel gear model",
+        "texts": [
+            " Marambedu [5] developed a full finite element model for the analysis of straight bevel gears for the prediction of load distribution, transmission error and stress including also the effect of surface deviations in the load distribution. Moreover, deterioration of one or several teeth affects the mesh stiffness and consequently the dynamic behavior of the transmission. For spur gear, Chaari et al. [2] quantified the reduction in gearmesh stiffness according to the severity of the damage. In this paper a lumped parameters model of a single stage bevel gear transmission is presented. The influence of tooth crack is investigated. The proposed dynamic model of a spiral bevel gear pair is shown in Fig. 1, it is divided into two rigid blocks. Each block has five degrees of freedom (three translations xi ; yi ; zi and two rotations m and 1 for pinion and 2 and r for wheel). The two gear bodies are considered as rigid cone disks and the shafts with torsional stiffness. The mesh is modeled by a linear stiffness ke.t/ acting along the line of action, this stiffness will be discussed later. Geometrical errors are modeled in the model by adding a transmission error e.t/ as a displacement on the line of action of the bevel gear gearmesh"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002774_s12206-011-1203-4-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002774_s12206-011-1203-4-Figure2-1.png",
+        "caption": "Fig. 2. Single bolt connection in the slewing bearing.",
+        "texts": [
+            " This paper will consider three major factors including preload, friction coefficient of the contact surface between the bolt and connecting part, and the height of the bushes. The bolt stress caused by combinations of different levels of the factors is computed with FEM. In view of the complexities involved in overall bolt connection performance, this paper will focus on single bolt connection performance since analyzing bolt connection performance by considering only the stress limit is common practice and produces highly reliable results. The model of a single connecting bolt is shown in Fig. 2. As mentioned above, three main factors for the bolt connection performance of slewing bearings will be analyzed with the orthogonal design method. According to this method, three factors are considered as independent variables, and the effect on the bolt connection performance is studied. There are three factors for the orthogonal design, and each factor is divided into five levels, as shown in Table 1. \u03c3s is the yield limit stress value of the bolt material. It needs to be computed 125 times with the full factors orthogonal design method, which will result in the establishment of 125 models for the bolt connection"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001078_ecce.2009.5316538-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001078_ecce.2009.5316538-Figure9-1.png",
+        "caption": "Fig. 9: Iron loss density in the induction motor prototype.",
+        "texts": [
+            " The stator tooth tips represent a special region where both radial and tangential flux-density components have relevant values (Fig. 8). Particularly at and above rated voltage, this problem is very challenging because of the extremely high value of the flux density. 3. Fourier series decomposition is applied for all the fluxdensity waveforms The first 11th harmonic orders are considered for practical reasons. 4. Equations (18-20) are applied to compute the iron losses in each mesh element of the iron regions. 5. Summation of all the losses per element gives the total iron loss under sinusoidal voltage supply (15). Fig. 9 shows the iron loss distribution under sinusoidal supply voltage conditions. The highest iron loss density is predicted in the stator teeth region. 6. Apply relation (12) to account for PWM voltage supply. The tests with PWM supply have been performed for a fundamental 50 Hz and a switching frequency of 2 kHz. The tests have been performed with variable modulation index (fixed DC bus voltage) and with variable DC bus voltage (fixed modulation index). In all the tests a sinusoidal modulation waveform was adopted"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000280_s11249-006-9137-1-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000280_s11249-006-9137-1-Figure2-1.png",
+        "caption": "Figure 2. Rolling four-ball tribosystem: (a) photograph and (b) scheme; P\u2014applied load, n\u2014rotational speed, 1\u2014top ball, 2\u2014bottom balls, 3\u2014race.",
+        "texts": [
+            " Pitting depends on many various factors: material properties, surface machining, geometry of the tribosystem, working conditions and chemistry of lubricating additives in oils. Out of these factors an influence of the chemistry of antiwear (AW) and extreme-pressure (EP) additives on pitting needs more investigation for the reason of a lack of a complex, chemomechanical approach to explain phenomena occurring in rolling contacts and affecting the surface fatigue. The tests were performed using a rolling four-ball tribosystem. The three lower balls were free to rotate in a special race, driven by the top ball (figure 2). The lower balls were immersed in the tested oil. The antifatigue properties were characterized by the 10% fatigue life, denoted as L10, determined according to IP 300/82 [4]. L10 represents the life at which 10% of a large number of test balls, lubricated with the tested lubricant, would be expected to have failed. Test conditions were as follows: rotational speed 1450 rpm; applied load 5886 N (600 kgf); run duration\u2014until pitting occurs; number of runs 24 (accepted only those for which a fatigue failure occurred on the top ball)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003092_1350650112439659-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003092_1350650112439659-Figure1-1.png",
+        "caption": "Figure 1. The structural diagram of spiral oil wedge hydrodynamic bearing.",
+        "texts": [
+            " The critical shear stress of oil at EMP journal bearing surfaces was 1.12694 kPa, which was obtained by Jin et al.20 So the study on the influence of critical shear stress on wall slip is very necessary. Most slip analysis is based on 1D circumferential direction slip, but the wall slip analysis of 2D and complex flow is seldom. The values of critical shear stress are also different if work environment, friction materials, and lubricants are different. In this article, spiral oil wedge hybrid sleeve bearing (Figure 1) is studied, its outstanding features of the bearing are as follows: three arc oil recesses on whole circumference are titled and the bearing has oil feed holes and oil return holes at both ends of each recess. So the different structures of this new bearing make its static and dynamic characteristics and slip properties different from conventional bearings. By means of the finite differential methods, the influence of critical shear stress on slip characteristics of 2D journal bearing considering the circumferential and axial wall slip is studied",
+            " Since the direction of slip velocity cannot be known in advance, the front of scalar usb, wsb, usa, and wsa multiplies the corresponding coefficient culx, culz, cdlx, and cdlz when the formula is derived. The regulation of positive sign for fluid shear stress is shown in Figure 5(a), negative sign is in the opposite direction. If yx and yz are at PENNSYLVANIA STATE UNIV on May 22, 2015pij.sagepub.comDownloaded from greater than 0, culx and culz are 1. If yx, yz is less than 0, culx and culz are 1. Because the surface of the spiral oil wedge hybrid sleeve bearing consists of two parts, the common surface same as conventional bearing and the circular recess surface (Figure 1) caused by three oil recesses, the oil film thickness must be calculated, respectively. Film thickness at the common cylinder surface is written as h \u00bc c\u00fe e cos\u00f0 \u00de \u00f016\u00de Film thickness at the eccentric arc surface is written as h \u00bc c\u00fe e cos\u00f0 \u00de \u00fe R1 cos \u00fe e1 cos\u00f0 \u00de R cos \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 \u00bde1 sin\u00f0 \u00deR1 2 q 8< : \u00f017\u00de where e is the eccentricity, the position angle of the eccentric arc, \u00bc ztg =R, the attitude angle, R1 the radius of the arc surface, e1 the eccentricity of the eccentric arc, and the angular coordinate, \u00bc x/R. At the regions of inlet holes 2 (Figure 1), the pressure is taken to be equal to the feed lubricant pressure pin; at both ends of the bearing and the region of outlet holes 1, the pressure is taken to be equal to the ambient pressure that is defined as being zero; at the point of film rupture, the pressure boundary condition accords with the Reynolds boundary condition, that is p \u00bc 0, @p @ \u00bc 0. In the same state and calculation conditions of Ma et al.,19 it is proved whether the numerical results of 2D spiral oil wedge bearing based on critical shear stress model are right"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002954_biorob.2012.6290682-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002954_biorob.2012.6290682-Figure4-1.png",
+        "caption": "Fig. 4 A straigh",
+        "texts": [],
+        "surrounding_texts": [
+            "In this paper, we focus on bowel peristalsis as a model for a mechanism that can transport fluids, such as sludge with little water. We developed a peristaltic pump based on the bowel mechanism by using an artificial rubber muscle, and confirmed its capabilities. In addition, we develop new tube to achieve a perfect close of the tube, and confirm the basic characteristics of the new tube. Finally, we measure the suction pressure and confirm the perfect close of the tube.\nI. INTRODUCTION pump that is capable of transporting high-viscosity fluids and solid\u2013liquid mixtures is required in various disasters such as flood that cause soil liquefaction, and in various industrial settings including sewage treatment plants and food processing. In addition, growing demand in the medical field. Moreover, it should be as easy as possible to install such a pump system. Turbine-type pumps, piston-type pumps and squeeze-type pumps are often used to transport high-viscosity fluids and solid\u2013liquid mixture fluids. However, these types of pumps have disadvantages. Turbine-type pumps cannot exert a high discharge pressure; piston-type pumps are usually considerably large because high pressure is needed to transport large quantities of fluid; squeeze-type pumps are a large and complex equipment. Furthermore, it can be difficult to arrange the bent pipes needed by these systems because of high friction between the fluid and the pipe walls. Hence, an innovative transport Manuscript received January 28, 2012\nKunihiro Saito, Yoshihiro Hirayama and Yoshiki Kimura are with the Department of Precision Mechanics, Faculty of Science and Engineering, Chuo University, 1-13-27 Kasuga, Bunkyo-ku, Tokyo 112-8551, Japan (e-mail: k_saito@bio.mech.chuo-u.ac.jp)\nTaro Nakamura is with Department of Precision Mechanics, Faculty of Science and Engineering, Chuo University, 1-13-27 Kasuga, Bunkyo-ku, Tokyo 112-8551, Japan (e-mail: nakamura@mech.chuo-u.ac.jp)\nsystem is desired. Therefore, various peristaltic pumps have been proposed [1]-[4] to transport these fluids. Mangan proposed a peristaltic gripping device [3]. This device was inspired by the preying activity of aplysia californica, which is a type of ocean creature. However, this device cannot transport fluid, although they can transport solids, such as a small ball. Miki was prepared an artificial esophagus using a shape memory alloy [4]. However, this device is very small, and it is difficult to be larger. For this reason, usage environment is limited. Therefore, we focus on bowel peristalsis as a model for a mechanism that can transport fluids, such as sludge that contains little water. Thus, we developed a peristaltic pump based on bowel peristalsis by using an artificial rubber muscle driven by pneumatic pressure [5]. In a previous study, we have successfully demonstrated transport of high-viscosity fluids (viscosity of 19000 mPa s) and solid\u2013liquid mixtures (solid content rate of 50%) [6]-[8].\nHowever, the cylindrical tube that is currently being used is not completely closed. Therefore, reverse of transport fluid was happening, and reverse of only fluid when transport solid-liquid mixtures. Thus, in this paper, we develop a new tube to achieve perfect closure. We also aim at determining the basic characteristics of the new tube. This includes measuring the suction pressure and confirming the perfect tube closure.\nII. BOWEL PERISTALSIS Bowel peristalsis is a motion that uses circular muscles located in the intestinal tract wall. This motion propagates toward the anus from the mouth [9]. Fig. 1 shows the motion resulting from bowel peristalsis. Bowel peristalsis can be described by the following three processes:\nA\n978-1-4577-1200-5/12/$26.00 \u00a92012 IEEE 949",
+            "t\nc\nm\nc\nc\nd t A c p\nilation defor he cylindrical s seen from ylindrical tub ush out towa\nFlang\nAir ve\nPressureless\nPressurized\ns how the app\nessurized. Bo\ne expand du\nds only in th\nthe axial dir\nfour guide tr\nmation when tube is divide the axial e takes the s rd the center.\nFig. 2 Cr\ne\nnt\nArtificial muscle\nRadial direction\nContra\nFig. 3 Appear\nearance of the th the artific ring pressuriz e radial dire ection. The enches. These pressurizing d into four pa direction, up hape of four\noss section of a u\nRubber tube\nAxia directi\nction\nance of pressuriza\nunit changes ial muscle an ation; the art ction, and th cylindrical tu trenches trig the unit. Ther rts by these tre on expansion quarter circle\nnit\nChamber\nl on tion\nwhen d the ificial e unit be is ger a efore, nches. , the s that\nB. F artif natu mus dire fibe duri pres carb mus at th\nC. F (old guid Tab natu carb unit\nT tren pres pow app perp tube cyli\nh\nd\nTAB SPECIFICATION\nStraight-Fibe ig. 4 shows a icial muscle ral rubber lat cle has the c ction of the tu rs, and spread ng pressuriz surized, it exp on fibers do n cle contracts is time serves\nCylindrical T ig.5 shows the tube), Fig.6 e trenches (ne le I. These cy ral rubber lat on fibers that . he new cylind ches which surizing the er that acts lied to the endicularly t begins to b nder side expa\nLength [mm Diameter [mm\nExcrescenc eight [mm] angl\nlength [mm\nGuide trenc epth [mm] angle\nlength [mm\nThickness [m\nLE I S OF NEW UNIT\nr-Type Artific straight-fiberconsists of a ex and a carbo arbon sheet a be. The carbo s across its w ation. When ands only in t ot readily exp in the axial di as an actuato\nube cylindrical tu\nshows the c w tube), and lindrical tub ex. The old become cons\nrical tube is c trigger a unit. Fig.7 sh on the cylind\noutside of o an action si end from the nds.\n] ]\ne e [deg.] ]\n3\nh [deg.] ]\n1.5\nm]\nial Muscle type artificial tube made o n roving shee rranged para n sheet is mad idth as rubbe the artific he radial direc and axially; th rection. The c r.\nt fiber muscle\nbe that includ ylindrical tub its specificatio es are made o cylindrical tub trained when\nomposed of eq dilation def ows the pat rical tube. T the cylindr de. Therefore guide trench\n90 60\n80 2 0\n90 65\n1.5\nmuscle [7]. T f low-ammon t. This artific llel to the ax e of thin carb r layers expa ial muscle tion because t us, the artific ontractive for\nes carbon fib e composed ns are shown f low-ammon e includes fo pressurizing\nuiangular gu ormation wh tern diagram he air pressu ical tube a , the cylindri . After that, t\nhe ia ial ial on nd is he ial ce\ners of in ia ur the\nide en of re cts cal he",
+            "m w t i s g d c o t\nw o m\nA r i i s p t a s c r s n t b\nF\nF\nThe optimum odel shown hen the tube\nhe shape of t ndicates the hows the sha uide trench irection on th ompletely, th ther words, th han its initial\nhere n is the f the cylindri inimum inte\ns the numbe ubber expans ncreases. M ncreases, the ignificant be rogression an ends to incre nd n should b tudy. The c ylindrical tub ubber part be hould close c o fold lines s renches. Fig.9 y guide trenc\nig.6 New rubber\nig.7 Pressure acti\nnumber of g in Fig. 8. Th\nexpands unif he rubber par shape during pe in the final arrangement, e circumferen ere can be no e final length length. This c\nnumber of gu cal tube. In ( gral value n is\nr of guide tr ion required t oreover, as interference o cause of th gle. Thus, we ase as the num e as small as ondition for d e is that no tween guide t ompletely. Fir hould occur shows a sch hes.\nGuid\nExcresce\ntube and cross-se\nng on the cylindr\nuide trenches e figure pres ormly. The da t in its initial the expansio state. The gre which is in ce. To close t slack in the of the rubber ondition is ex\nide trenches a 1), n is an int derived as\nenches increa o reach the fi the number f each expand e reduction assume that t ber of guide possible. We w etermining th fold lines sh renches, and t st, we consid on the rubber ematic of the\nPressure\ne trench\nnce\nction view\nical surface\nis considered ents an axial shed line repr state, the sol n, and the re en circle show a countercloc he inside of th cylindrical tu part must be pressed as\nnd d is the dia eger. Therefo\nses, the amo nal expanded of guide tre ing part can b of the expa he required pr trenches incr\nill set n to 4 e axial length ould occur o he inside of th er the conditio part between rubber part d\nin the view esents id line d line\ns the kwise e unit be. In longer\nmeter re, the\nunt of shape nches ecome nsion essure eases, in this of the n the e unit n that\nguide ivided\ncompartmental"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000889_978-1-84882-694-6_13-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000889_978-1-84882-694-6_13-Figure3-1.png",
+        "caption": "Figure 3: Compact MRF Brake (CMRFC)",
+        "texts": [],
+        "surrounding_texts": [
+            "Locomotion is a most important skill of the activity of daily living (ADL) for human beings. Therefore gait trainings have been made a high priority in rehabilitative trainings. Normal gait is cyclic and can be characterized by timing of foot contact with the ground; an entire sequence of functions by one limb is identified as a gait cycle [1], [2], [3] as shown in Fig.1. Each gait cycle has two basic components: \u201cstance phase,\u201d which designates the duration of foot contact with the ground, and \u201cswing phase,\u201d the period during which the foot is in the air for the purpose of limb advancement. The swing phase can be further divided into three functional subphases: \u201cinitial swing,\u201d \u201cmid swing\u201d and \u201cterminal swing.\u201d In the same manner, the stance phase can be partitioned into five functional subphases: \u201cinitial contact,\u201d \u201cloading response,\u201d \u201cmidstance,\u201d \u201cterminal stance\u201d and \u201cpreswing\u201d [1], [4], [5]. In the normal gait, initial contact becomes \u201cheel-contact (or heelstrike),\u201d appropriately. And in the initial swing (or \u201ctoe-off\u201d), normal subjects can keep appropriate clearance from the\nground to prevent a malfunctional interaction with the ground.\nFor patients who have dysfunction of ankles, for example polio and peroneal nerve palsy, it is difficult to control their ankle by themselves. This problem causes \u201cdrop-foot\u201d or a lack of ankle dorsiflexion during the swing phase. In many case, they can not prevent from stumbling their toe with even small steps on the ground. Additionally, they have a tendency to incline their upper body more than healthy persons because of the rough motion to prevent from stumbling of the toe. It causes undesired energy-loss in walking.\nAn orthosis is defined as a device attached or applied to the external surface of the body to improve functions, restrict or enforce, or support a body segment [1], [6]. In order to improve the gait of the patients, lower limb orthoses are applied to them. In order to assist their ankle function, anklefoot orthoses (AFOs) are often used to restrict their involuntary plantarflexion and so on.\nService Robotics and Mechatronics",
+            "From a long time ago, many researches focusing on powered dynamic-controllable lower extremity prostheses with robotics technologies are reported [7]-[11] and some kind of products are released (Otto Bock C-Leg [12], Ossur RHEO KNEE [13]). These prostheses are designed to be patient-adaptive or environment-adaptive. These reports give us grate amount of useful information about gait control [14] for the idea of powered and dynamic-controllable AFOs.\nActually, powered AFOs are focused on in recent years. Some kinds of the powered AFO were reported [15] by using several types of actuators, e.g. pneumatic actuation system (Keith E. Gordon and Daniel P. Ferris [16]), ball screw drive system (Bashir M. Y. Nouri and Arafat Zaidan [17]), series elastic actuator (Joaquin A. Blaya and Hugh Herr [18]), and so on. However, developments of such a powered AFO device are more difficult than that of the powered prostheses because we have to consider dynamics of a paralyzed leg or foot (inertia, viscosity, elasticity and voluntary/involuntary force from them). Additionally, weight saving is required strongly.\nIn this study, we suggest passive controllable AFOs with compact brake devises for dynamic gait controls [19]. In particular, the control of drop-foot can be realized by only using passive devices. The passive controllable AFO also have a great advantage for cost, safety and downsizing.\nFor the beginning of this paper, we describe development of a compact MR fluid brake. Next a passive controllable AFO system with this brake is explained. Then, we discuss a control method. And finally experimental results of gait control for a polio patient are presented.\nMR fluids (MRFs) are kinds of functional fluid, which is composite materials of non-colloidal solution and magnetic metal particles (e.g., iron particles). The diameter of the particle is 1~10 micrometers. Its rheological properties can be controlled by applying a magnetic field [20]. The response of its viscous change is very rapid (about several milliseconds) [21], [22] and repeatable with large range. By utilizing the viscosity change of the MRFs with share flow mode, we can develop a brake device with good responsibility, higher torque/inertia ratio than any other conventional brakes, e.g. powder brakes and electromagnetic brakes [21]-[23]. In recent years, a prosthetic knee using the MR fluid braking device was reported [11] and has produced by a prosthesis maker [13].\nA conceptual drawing of a compact MRF brake (CMRFC) is shown in Fig.2. A coil is rolled round an output shaft and it generates the magnetic flux shown by dashed lines in the drawing. Multi-layered disks are fixed on the stator-parts and rotator-parts. The MRF is filled between these disks.\nAs you see in this figure, multi-layer structure is utilized for enlargement of the brake torque. However, we have to impress a magnetic flux density of about 0.5~1.0 tesla on the MRF layers in order to make use of the viscous change of the MRF effectively. Because the MRF is a material which has a\nlow magnetic permeability, it is important to reduce the total gap of the MRF layers for the reduction of electric power consumption.\nBased on this restriction of the gap-size, total thickness of the MRF layers was decided to be less than 0.5~1.0 mm, empirically. Therefore, we suggest to utilize narrow-gaps of 10~100\u03bcm for each gap. The gap-size, the diameter of disks, the number of the layered disks and the number of turns of the magnetic coil should be estimated based on results of magnetic analyses and processing (or assembling) accuracy of the multi-layered disks.\nWe formulated the way to estimate output torques of the CMRFC. The analysis flow is shown as follows;\n(1) Geometric design of a CMRFC with 3-D CAD software,\n(2) Magnetostatic analysis for the estimation of the magnetic",
+            "flux density with CAE (FEM) software,\n(3) Decision of a yield stress of the MRF depending on the results of the process (2),\n(4) Calculation of a maximum torque depending on the results of the process (3) and other size-parameters.\nIn the process (2), we need to input nonlinearity data (B-H curve) of ferromagnetic materials (silicon steal etc.) and the MRF. Commercially produced MRF (140CG, Lord Corp.) was used as an MRF in this study.\nIn the process (3), we referred to the characteristics data between the yield stress and the magnetic flux density of 140CG, which is presented by Lord Corp. [24].\nFigures 3 show a developed compact CMRFC. Specification data of this clutch is shown in table 1. Multi-layered disks are fixed with gaps of 50\u03bcm accurately. The MRF (140CG, Lord Corp.) is filled between these gaps completely. Figure 4 shows a cross-sectional view of the CMRFC. In this figure, black broken line means a loop of the magnetic flux. The condition of filled fluid can be checked through the inlet hole made direct above multi-layered disks as shown in the sectional view of Fig.4. The number of turns of the coil is 191 turns. According to the result of a magnetostatic analysis mentioned above, application of the electric current of 1A to the coil generates magnetic flux density of 0.55~0.65 tesla in the MRF layers.\nTorque characteristics of this device were measured by an experiment setup as shown in the Fig.5. The casing of MR brake was fixed on a turning table and rotated at a constant\nspeed of 1 rad/s. A lever-arm of which length is 250 mm was fixed on the output shaft of the brake. We measured braking torque with a force sensor connected with this arm.\nAt first, static torque tests were conducted at constant speed of 1 rad/s. The experimental results are shown in Fig.6. White circles are experimental data. Black squares mean the results of the analysis mentioned in the previous session. As shown in this figure, the experimental result under the static condition indicates good similarity to the analytic result. Therefore, we can predict the maximum torque of the CMRFC with the suggested method.\nSecondly, we also conducted dynamic torque tests (stepresponse). Time constant (63% response) of step-response was about 20 milliseconds as shown in Fig.7. This results"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002482_j.triboint.2009.12.042-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002482_j.triboint.2009.12.042-Figure1-1.png",
+        "caption": "Fig. 1. Surface texture primitives.",
+        "texts": [
+            " Therefore there should be a relationship between the shape of asperities and the tribological behaviour. On the worn surfaces of the ceramic balls pit-like structures caused by brittle fracture occur on the surface. These pits may also relate to the wear and normal load. ll rights reserved. Q. Hao). A feature based characterization technique was published, which characterizes the surfaces according to their texture primitive [8]. This means, a surface can be divided into regions consisting of hills and regions consisting of valleys, like a landscape shown in Fig. 1. These regions consist of critical points (peaks, saddle points and pits). The boundaries between the hill regions are the course lines and the boundaries between the valley regions are the watershed lines. By using a Wolf pruning method [9] the information not relevant to the application can be removed. Then the surface is segmented according to the hill or valley regions. A method for detecting particle and pore structures on a surface was also developed to characterize an AlSi cylinder liner surface [10]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001077_detc2009-87092-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001077_detc2009-87092-Figure1-1.png",
+        "caption": "Figure 1. Two-stage cycloid drive [2].",
+        "texts": [
+            " In comparison with conventional gear mechanisms, cycloid drives have very high precision and small size, making them attractive for limited space applications, such as robots, machine tools, linear axis positioning systems and a wide variety of factory automation equipment, especially where precise positioning, stiffness, and shock-load absorbing capability are required [1]. The two-stage cycloid drive has been developed to achieve a high reduction ratio in limited space [2]. It consists of input spur gear and output cycloid drive stages (input eccentric shaft, cycloid disk, pin-rollers, and output disk), as shown in Figure 1. Two or more eccentric shafts transmit not only torque of the first spur gear stage to the cycloid disk but also that of the cycloid disk to the output disk. Therefore, bearings mounted in the eccentric shaft are expected to have a significant effect on performance of the two-stage cycloid drive. Although there have been previous studies on contact force and stress analysis [3-5], these investigations did not consider finite bearing stiffness as well as nonlinear Hertz contact stiffness that depends on contact geometry and contact force"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001766_ijnt.2009.022922-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001766_ijnt.2009.022922-Figure12-1.png",
+        "caption": "Figure 12 (a) Chemical reduction of polypyrrole followed by its in situ oxidation and consequent reduction of Ag+ to Ag0 and (b) in situ reduction of Ag+ to Ag0 by polyaniline",
+        "texts": [
+            " Polyaniline forms in the emeraldine state when ferric chloride or ammonium persulfate is used as the oxidant. The electrochemical potential of the oxidation of emeraldine to pernigraniline can couple directly with the reduction potential for silver ions and similarly deposit metallic silver on the conducting polymer surface. The silver can form either as a continuous coating or as discrete silver nanoparticles on the surface of the polymer at lower Ag+ concentrations. The reaction schemes are shown in Figure 12. SEM images show the silver deposited from a AgNO3 solution over a longer time on a polypyrrole-paper hybrid reduced with hydrazine, is present as a continuous silver coloured coating (Figure 13(a)). For polypyrrole-cellulose fibres and polypyrrole-kaolinite particles that have been reduced with sodium borohydride and immersed in AgNO3 solution, the silver is present as discrete nanoparticles about 100 nm in size (Figure 13(b) and (c)). As these do not collectively scatter light, the material retains the dark blue-black colour of the polymer"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002532_detc2012-70485-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002532_detc2012-70485-Figure2-1.png",
+        "caption": "FIGURE 2. Omnicopter Control Methods. M1: Fixed ducted fan angles with varying speeds (left); M2: Variable ducted fan angles and variable speeds (right)",
+        "texts": [
+            " It has five propellers: two fixed major coaxial counter-rotating propellers in the center used to provide most of the thrust and adjust the yaw angle, and three adjustable angle small ducted fans located in three places surrounding the airframe to control its roll and pitch. Unlike quadrotor or typical trirotor MAVs [2, 3], the Omnicopter has different motion principles and control modes. Increasing or decreasing the five propeller\u2019s speeds together generates vertical motion. The yaw movement results from different speeds of the two counter-rotating coaxial propellers. The roll and pitch motions can be generated using two methods (M1 and M2), as shown in Fig. 2. For M1, fixed ducted fan angles with varying fan speeds are used; and for M2, varying both the angles and speeds of the ducted fans are employed for attitude control. With method M1, the difference between the speeds of Fan 4 and 5 produces roll motion coupled with lateral motion. The pitch rotation and the corresponding lateral motion result from the difference between Fan 3\u2019s speed and the collective effect of rotation of Fan 4 and 5. The second control method (M2) is to adjust the angles of the surrounding ducted fans, with the fan\u2019s speeds variable or fixed, to generate the roll and pitch motions",
+            " The potential energy is: V =\u2212mg\u03b53 Based on the kinetic and potential energy equations, we can derive the equations of motion using the Euler-Lagrange formalism [16] \u0393 j = d dt ( \u2202T \u2202 q\u0307 j )\u2212 \u2202T \u2202q j + \u2202V \u2202q j where q j denotes a component of the generalized coordinates vector, j = 1, 2, \u00b7 \u00b7 \u00b7, 6. Finally, the equations of motion are obtained as the following: m\u03b5\u03081 = F\u03b51 m\u03b5\u03082 = F\u03b52 m\u03b5\u03083\u2212mg = F\u03b53 Ixx\u03c9\u0307x\u2212 (Iyy\u2212 Izz)\u03c9y\u03c9z = \u03c4x Iyy\u03c9\u0307y\u2212 (Izz\u2212 Ixx)\u03c9x\u03c9z = \u03c4y Izz\u03c9\u0307z\u2212 (Ixx\u2212 Iyy)\u03c9x\u03c9y = \u03c4z (6) In Eq. (6), F\u03b51 , F\u03b52 and F\u03b53 represent the generalized forces of the translational movement in Eq. (2) for a complete system model. In this paper, we focus on an Omnicopter being controlled with method M1 (Fig.2(left)): fixed angle ducted fans with varying speeds. We assume that the surrounding ducted fans point vertically upwards, thus the angles \u03b1 , \u03b2 and \u03b3 in Eq. (2) become 90\u25e6. In this case, the Bx and By components of F become 0, and the equations of motion of the translational subsystem become m\u03b5\u03081 =\u2212(c\u03c8s\u03b8c\u03c6 + s\u03c8s\u03c6)U1 m\u03b5\u03082 =\u2212(s\u03c8s\u03b8c\u03c6 \u2212 c\u03c8s\u03c6)U1 m\u03b5\u03083\u2212mg =\u2212c\u03b8c\u03c6U1 where U1 = kF1 (\u21262 1 + \u21262 2) + kF2 (\u21262 3 + \u21262 4 + \u21262 5). Now similar in dynamics to the quadrotor MAV, as shown in [5], the linear equations of motion are simple in the inertial reference frame, while the angular equations are advantageous to be expressed in the body-fixed coordinate frame"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003135_iros.2011.6095094-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003135_iros.2011.6095094-Figure4-1.png",
+        "caption": "Fig. 4. CAD rendering of biorobotic fish with mass distribution cavities circled and pectoral fin module boxed",
+        "texts": [
+            " The dorsal and anal fins are attached to the tail section, which also houses these fins\u2019 actuators and controllers, and additional weight compartments. The tail also houses the caudal fin\u2019s actuators and controllers, and is currently rigid, but can be made flexible so that it can be actuated to undulate like the biological fish. The peduncle contains the base for caudal fin (Fig. 3, D). The locations of the robot\u2019s center of mass (CM) and center of buoyancy (CB) can be adjusted, by distributing weights throughout cavities within the individual body sections (Fig. 4), to tune its stability and buoyancy. The robot is currently configured with the CM below the CB, such that the body is statically stable and neutrally-buoyant. As our understanding of body dynamics improves, the distance between the CM and the CB can be reduced and eventually inverted, which will make the system want to flip upside down. This ability will be used to explore the interdependencies of dynamic instability of the body and its maneuverability/controllability. Sensors (described in Section IV) are placed on the body surface and internally to measure changes body orientation and the surrounding fluid flow",
+            " Subsequent generations of the robotic fish will incorporate multi-DOF actuated fins as described in [2, 4], that can emulate multiple gait patterns observed in the sunfish. Each fin-ray is actuated by low-friction, stainless steel wire rope routed through a length of flexible nylon conduit in a pull-pull configuration. The fixed length of conduit allows for flexibility in the relative placement of each fin ray and its associated servomotor while maintaining wire tension, which is necessary when considering a flexible body where the distances and positions of the fins and servomotors vary during movement. Each fin is driven by an actuator module (Fig. 4) that includes servomotors and a microcontroller within a waterproof enclosure, with the wire tendons connecting the servomotors to the fin bases (Fig. 3, A). Fin microcontrollers and a central microcontroller communicate via a CAN bus. This common architecture was devised so that multiple configurations of fins could be introduced easily into the robotic system and to mimic the distributed neurological architecture of its biological counterpart. The distributed controllers also enable different configurations of neural controllers to be implemented as our understanding of the sunfish neuro-anatomy develops"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000093_s11431-006-0238-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000093_s11431-006-0238-8-Figure1-1.png",
+        "caption": "Fig. 1. (a) Schematic representation of the PLDM process; (b), (c) the equipment photos.",
+        "texts": [
+            " The characteristics of the PDM technology are applied, and the better collective laser beam as a supplemental energy resource is adopted to make the high temperature material prototype with high quality. The effects of laser beam parameters, such as average power, pulse width, repetition frequency and the angle between laser beam and plasma arc beam, on the performance characteristics of arc beam and shaping characteristic in the process of hybrid plasma and laser deposition manufacturing are studied experimentally and are confirmed during the actual single-layer or multi-layer deposition shaping. The principle of PLDM technology is shown in Fig. 1. The shape, dimension and function of the conception model are gained by means of concept design, and then work out a CAD model with the help of a shaping software. According to the requirement of the component\u2019s dimensional accuracy, the CAD model is dealt with slices and transferred into an STL document. In accordance with the shape of every layer, the control system draws up to route, generates the CNC code to control the movement of the NC machine. On the one side of the plasma arc, the laser beam is focused on the plasma arc into the molten pool through adjusting the optical path for light transmitting and focusing",
+            " The high-speed CCD camera was employed to obtain the picture around the plasma arc. And then the clear sketch of the picture was gotten after dealing with the plasma arc picture. By means of the obtained picture sketch, the diameter of arc was measured and the influence of the laser technological process parameter on the plasma arc column shape was analyzed. The diameter of the arc was selected at the same distance below the plasma nozzle. The PLDM equipment with five-axis numerical controlling function as shown in Fig. 1(b) was used in the experiment. The range of the laser average power was from 4 W to 280 W, the transmission and focus of laser could be adjusted freely in the optical path system. (i) The influence of the laser average power. Fig. 2(a) shows the relationship between plasma arc characteristic and laser average power. Fig. 2(b) is the plasma arc sketch dealt with by Gauss-Laplace method. The relationship between average power and plasma arc diameter is shown in Fig. 3. It is clear that the plasma arc is becoming smaller when the laser average power is increasing, which means that the laser has the compression effect on the plasma arc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001680_elan.200804612-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001680_elan.200804612-Figure1-1.png",
+        "caption": "Fig. 1. Versatile flow head for microsensors testing, a) crosssection of the flow-head; b) inplant design cross-section; c) top view of the implant.",
+        "texts": [
+            " At first, a flow head for ten sensors was manufactured from a block of transparent plexiglass (f\u00bc 65 mm and 20 mm thick). The transparency of the constructing material is very useful for the control of air bubbles and flow properties inside the head, especially during the head geometry optimization. Plexiglass has been chosen as a testing-head material also because of its good mechanical properties and resistance to aqueous solutions of electrolytes. The cross section of the designed head is presented in Figure 1a. The measured sample transported by a peristaltic pump reaches a chamber (95 mL), where it divides into ten streams (26 mm long and f\u00bc 1,5 mm). It is assumed, that the streams are equal in the flow, which results from the symmetry of the drilled channels and the chamber. At the ends of the channels there are sensor implants, which were made from separate blocks of polymer and glued. The design and the top view of the implant with mounted backside contact microsensor are presented in Figure 1b and 1c, respectively. The idea of the implant construction allows easily and reproducibly mounting in it various types of microsensors (SSE; Ion-Selective FET: ISFET; Chemical FET: CHEMFET; optrode) [13 \u2013 15]. The microsensor chip and the sensor housing is in a form of square 5 5 mm so the square tip of the plug fits to the implant cavity and there is no possibility of twisting it on the back surface of the chip, what could damage both electric pads or a sensing membrane on the opposite side of the mounted sensor"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001435_acc.2007.4282143-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001435_acc.2007.4282143-Figure4-1.png",
+        "caption": "Fig. 4. Robot navigation for different values of B",
+        "texts": [
+            " The navigation parameters are calculated online to navigate the robot towards the intermediary goals. These goals are chosen based on the sensory data. in [23] Q\u2013learning is used to fix the navigation parameters. We restrict ourselves to stationary. However, moving obstacles can be considered too. This task can be accomplished by considering relative kinematics equations between the robot and the obstacles. V. SIMULATION We present a few simulation examples in the absence and presence of obstacles. In the example of figure 4, the robot initial configuration is given by (20, 20, \u03c0). The aim is to drive the robot to a final configuration given by ( 20, 20, \u03c0 4 ) . As shown in figure 4, this task is accomplished successfully by using linear navigation functions for different values of B. The scenario of figure 5 is similar but with a final configuration ( 20, 20, \u03c0 2 ) . It is clear that different values of B result in different paths. The scenarios of figures 6 and 7 show navigation in the presence of obstacles. The robot\u2019s initial configuration is (20, 20, \u03c0), and the final configuration is ( 20, 20, \u03c0 2 ) . Two different paths are accomplished by the robot by using different values of B"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002698_s1068798x10040234-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002698_s1068798x10040234-Figure2-1.png",
+        "caption": "Fig. 2. Simulation of the adaptive technological control system in ensuring the specified physicomechanical parameter: maximum tensile stress \u03c3 in the machined sur face: RC, rated characteristic \u03c3 = f(s); s is the supply.",
+        "texts": [
+            " Consider the simulation of the adaptive technolog ical control system in ensuring the specified physico mechanical parameter: the maximum tensile stress \u03c3 in the machined surface during external turning. Con trol by Eqs. (6)\u2013(8) is based on the mathematical model [3] MPa. This model is reduced to the form in Eq. (5), by set ting constant \u03c1 where \u03c1 = 0.5 mm is the rounding radius of the cutting edge. We assume that maximum tensile strength must be ensured in the machined surface: \u03c3_zad = 250 MPa, with tolerance \u0394 = \u00b110%. To ensure the specified \u03c3, we need s = 0.177 mm/turn, v = 120 m/min, and t = 1 mm. This corresponds to point 1 on the rated char acteristic in Fig. 2. As a result of external perturbations, \u03c3 passes beyond the tolerances. The calculated value \u03c3_c may be either larger than (\u03c3_zad + \u0394\u03c3_zad) or less than (\u03c3_zad \u2013 \u0394\u03c3_zad); this corresponds, respectively, to points 2 and 3 in Fig. 2. In other words, the rated char acteristic may be shifted up or down by the external perturbations (points 2 and 3 in Fig. 2). We now consider the case where \u03c3_c = 216.5 MPa (corresponding to point 3 in Fig. 2) is less than (\u03c3_zad \u2013 \u0394\u03c3_zad). The corrected supply to obtain the specified parameter \u03c3_zad is determined on the v const= t const= FMPc FMP_zad \u0394FMP_zad+\u2265 FMPc FMP_zad \u0394FMP_zad\u2013\u2264 CTfmp FMPc tzpxf ztyf+ = t FMP( ) FMP_zad CTfmp \u239d \u23a0 \u239b \u239e 1 zpxf ztyf+ = \u23ad \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23ac \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23ab , \u03c3 285v0.26s0.38\u03c10.24 ,= \u03c3 constv 0.26s0.38 ; const 285\u03c10.24 ,= = basis of Eq. (7) Thus, to ensure \u03c3_zad = 250 MPa (in machining conditions that are varying randomly), we need to increase the supply to s = 0.258 mm/turn (point 3 ' in Fig. 2). After correction, we obtain \u03c3_c = 244.5 MPa. The error when using Eq. (7) is 250 \u2013 244.5 = 5.5 MPa, which is 2% of the specified value. Thus, \u03c3_c is within the tolerances. When the calculated value \u03c3_c = 281.5 MPa (point 2 in Fig. 2) is larger than (\u03c3_zad + \u0394\u03c3_zad), the cor rected supply (to ensure that \u03c3_zad = 250 MPa) is s = 0.129 mm/turn. After correction, \u03c3_c = 245.6 MPa. The error when using Eq. (7) is 250 \u2013 245.6 = 4.4 MPa or 1.7% of the specified value; \u03c3_c is within the tolerances. Error in the control law may be due to the assump tion that the coefficients in the corrected control parameters (s, v, t) are constant and do not vary under the external perturbations. In the given examples, v 120= t 1= 216.5 250 0.1 250\u00d7\u2013\u2264 Cvfmp 216"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001933_s11071-010-9720-8-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001933_s11071-010-9720-8-Figure2-1.png",
+        "caption": "Fig. 2 Cross section of HSFD",
+        "texts": [
+            " The non-periodic behavior of this system is characterized using phase diagrams, power spectra, Poincar\u00e9 maps, bifurcation diagrams, Lyapunov exponents and the fractal dimension of the system. Figure 1 shows a rigid gear pair supported on a hybrid squeeze-film damper (HSFD) in parallel with retaining springs. The bearing consists of four hydrostatic chambers and four hydrodynamic regions. The oilfilm supporting force is dependent on the integrated action of hydrodynamic pressure and hydrostatic pressure of HSFD. Figure 2 represents the cross section of HSFD. The structure of this kind bearing should be popularized to consist of 2N (N = 2,3,4, . . .) hydrostatic chambers and 2N hydrodynamic regions. In this study, oil pressure distribution model in the HSFD is proposed to integrate the pressure distribution of dynamic pressure region and static pressure region as described in Sect. 2.1. Figure 3 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; 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; K is the stiffness coefficient of the gear mesh, C is the damping coefficient of the gear mesh, e is the static transmission error and varies as a function of time",
+            "1 The instant oil-film supporting force for HSFD To analyze the pressure distribution, the Reynolds equation for constant lubricant properties and noncompressibility should be assumed, then the Reynolds equation is as follows [17] 1 R2 \u2202 \u2202\u03b8 ( h3 \u2202p \u2202\u03b8 ) + 1 R2 \u2202 \u2202z ( h3 \u2202p \u2202z ) = \u221212\u03bc\u03a9 \u2202h \u2202\u03b8 + 12\u03bc \u2202h \u2202t . (1) The supporting region of HSFD should be divided into three regions: static pressure region, rotating direction dynamic pressure region and axial direction dynamic pressure region, as shown in Fig. 2. In the part of HSFD with \u2212a \u2264 z \u2264 a, the long bearing theory (see Appendix) is assumed and Reynolds equation is solved with the boundary condition of static pressure region pc,i acquiring the pressure distribution p0(\u03b8). In the part of HSFD with a \u2264 |z| \u2264 L 2 , the short bearing theory (see Appendix) is assumed and solves the Reynolds equation with the boundary condition of p(z, \u03b8)|z=\u00b1a = p0(\u03b8) and p(z, \u03b8)|z=\u00b1L/2 = 0, yielding the pressure distribution in axis direction dynamic pressure region p(z, \u03b8)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001741_17543371jset75-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001741_17543371jset75-Figure1-1.png",
+        "caption": "Fig. 1 The experimental apparatus used to project tennis balls at a Head Ti S6 racket",
+        "texts": [
+            " Tennis balls approved by the International Tennis Federation were fired from a ball projectile device and each impact was recorded at a high speed using two cameras, to enable a three-dimensional (3D) analysis to be carried out. Six independent variables were varied during the experiment: (a) the inbound ball velocities, measured on three axes; (b) the impact positions of the ball, measured in two directions across the racket face; (c) the level of restrictive torque about the racket axis. An impact volume was manufactured from Bosch Rexroth [7] aluminium extrusions (Fig. 1). A bank of lights illuminated the impact volume containing the racket and mount, and a racket mount was used to change the position and angle of the racket prior to impact and to provide the restrictive torque used to simulate grip. Two large sliding doors with clear polycarbonate windows gave access to the impact volume for setting up and calibration. The polycarbonate windows ensured an unobstructed view of impact while protecting the user from high-velocity balls. A BOLA ball projection device was used to fire the ball at the racket"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000516_s11044-008-9104-8-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000516_s11044-008-9104-8-Figure2-1.png",
+        "caption": "Fig. 2 A three link system",
+        "texts": [
+            " (64) For any point P on body I , suppose it has coordinates r(I ) P = [x(I) P , y (I) P , z (I) P ]T in the body fixed frame SI , then in frame SI we can write the velocity of the point as v(I ) P = I v+ I \u00b7r(I ) P = [ I3\u00d73 03\u00d71 ] I\u00b7V\u00b7 {\u2212r(I ) P 1 } (65) and acceleration a(I ) P = [ I3\u00d73 03\u00d71 ] I\u00b7A\u00b7 {\u2212r(I ) P 1 } + [\u2212I3\u00d73 03\u00d71 ] \u00b7 I V \u00b7 I V \u00b7 {\u2212r(I ) P 1 } = [ I3\u00d73 03\u00d71 ] \u00b7 ( I A\u2212 I V\u00b7 I V ) \u00b7 {\u2212r(I ) P 1 } . (66) We also have partial velocity \u2202v(I ) P \u2202\u03b3\u0307J,k(J ) = [ I3\u00d73 03\u00d71 ] \u00b7 \u2202 I V \u2202\u03b3\u0307J,k(J ) \u00b7 {\u2212r(I ) P 1 } = [ I3\u00d73 03\u00d71 ] \u00b7 Ad I R J (NJ,k(J ) ) \u00b7 {\u2212r(I ) P 1 } (67) and partial angular velocity \u2202\u03c9I \u2202\u03b3\u0307J,k(J ) = \u23a1 \u23a30 0 \u22121 0 0 0 0 0 1 0 0 0 \u23a4 \u23a6 \u00b7 \u2202 I V \u2202\u03b3\u0307J,k(J ) \u00b7 \u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 1 1 1 0 \u23ab\u23aa\u23aa\u23ac \u23aa\u23aa\u23ad+ \u23a1 \u23a31 0 0 0 1 0 0 0 1 0 0 0 \u23a4 \u23a6 \u00b7 \u2202 I V \u2202\u03b3\u0307J,k(J ) \u00b7 \u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 0 0 \u22121 0 \u23ab\u23aa\u23aa\u23ac \u23aa\u23aa\u23ad (68) for any J \u2208 L\u0302(I ), 1 \u2264 k(J ) \u2264 N(J). 5 Example A system of three uniform links is shown in Fig. 2. Link 1 has mass m1 and length L1 and is connected to the ground at O1 through a revolute joint; link 2 has mass m2 and length L2 and is connected to link 1 at O2 through a revolute joint; link 3 has mass m3 and length L3 and is connected to link 2 at O3 through a revolute joint. The mass centers of the three links are represented as G1,G2, and G3, respectively. The three links have body fixed frames O1xyz,O2xyz, and O3xyz, respectively. With respect to the corresponding body fixed frames and mass centers, link 1 has principal moment of inertia matrix JG1 = diag(J (1) 1 , J (1) 2 , J (1) 3 ), link 2 has principal moment of inertia matrix JG2 = diag(J (2) 1 , J (2) 2 , J (2) 3 ), link 3 has principal moment of inertia matrix JG3 = diag(J (3) 1 , J (3) 2 , J (3) 3 )"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003878_demped.2013.6645747-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003878_demped.2013.6645747-Figure6-1.png",
+        "caption": "Fig. 6. The healthy double cage induction motor.",
+        "texts": [
+            " Same like in the previous paragraph II, 4 different cases are investigated: a) The motor is healthy and the voltage supply is symmetrical, b) The motor is healthy and the voltage supply is asymmetrical, c) The motor has a broken bar and the voltage supply is symmetrical and d) The motor has a broken bar and the voltage supply is asymmetrical. When the voltage supply is symmetrical the applied voltage is 380V for all phases. In the asymmetrical supply condition the first and second phases have 380V, while the third has 353.5V. The simulated, healthy, double cage induction motor is presented in Fig. 6. In this section, the double cage induction motor operates for both healthy and faulty conditions at 1460rpm under symmetrical voltage supply. In Fig. 7-a, the line current frequency spectrum, for both cases, is presented. The results totally agree with previous work [17], since it is obvious that the fault signatures present much weaker amplitudes In Fig. 8-a the zero-sequence current spectrum is presented, for the healthy double-cage induction motor and the one with the broken bar fault. The strongest fault signature is located at 147"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure6-1.png",
+        "caption": "Figure 6. The transmission of motion from a rotating link to a counterweight.",
+        "texts": [
+            " For example, assume that for locating the counterweights we employ certain rotating links of the mechanism, whose laws of motion are known. In such a case it is necessary to place the known values of the sines and cosines of angles ~ and ~2 in equations (4a) and (5a). In order that the conditions of the equations be satisfied, it is sufficient to make the values Pl and P2 variable. Such a solution can be carried out in reality by attaching counterweights to selected links of the mechanism by means of sliding pairs, and by forcing them to roll along stationary cams of appropriate profiles, as shown in Fig. 6. 328 Let us now examine dynamic balancing of a mechanism, which follows after static balancing. We equate to zero the sum of the moments, relative to point O, of all the forces (Fig. l) 1/ t t p r , r ' t , t L, + C,,xZ~ + C,:x, + ~ mico x, zi+ ~ m~co :,x,.= 0, I | (7) where C,x and C= are the projections of the resultant vector on the coordinate axes; co~, and co':, are projections of the acceleration of the center of the i th counterweight on the coordinate axes. Equation (7) represents the only requirement which must be fulfilled"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002781_iccase.2011.5997721-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002781_iccase.2011.5997721-Figure4-1.png",
+        "caption": "Figure 4. Finite element model.",
+        "texts": [
+            " Hereinafter, the degree of dynamic eccentricity is defined as %100\u00d7\u0394= g ee (6) where g is the average airgap length, and wrOOe =\u0394 is the dynamic eccentricity vector in PMSG. Notice that this dynamic eccetricity vector is fixed in all angular positions of the PM rotor, however its angle varies as can be seen in Fig. 3(b). Thus, in such a condition, the air gap around the PM rotor is time varying and nonuniform. Here, the static and dynamic eccentricity is realized by placing the rotor in a new position with mesh regeneration in the FEM. Fig. 4 shows the corresponding FEM for the healthy PM generator that is used in this paper. The transient equations of the multi-loop circuits and the motion equations are combined with the magnetic field equations in the finite element modeling procedure. Thus, we can obtain the corresponding stator terminal currents with different kinds of inter-turn short circuit faults and different degree of dynamic eccentricity. III. HYBRID FAULT ANALYSIS It has been pointed out in [17] that the DE fault can be diagnosed by using a novel pattern frequency, which can be calculated as styeccentrici f P kf \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 \u239f \u23a0 \u239e \u239c \u239d \u239b \u2212\u00b1= 121 (7) where P is the number of pole pairs, k is an integer number 1, 2, 3\u2026, and fs is the supply frequency"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.9-1.png",
+        "caption": "Fig. 1.9. Simple car model with side force characteristics for front and rear (driven) axle.",
+        "texts": [
+            "4 will be treated first after which the simple model with two degrees of freedom is considered and analysed. This analysis comprises the steady-state response to steering input and the stability of the resulting motion. Also, the frequency response to steering fluctuations and external disturbances will be discussed, first for the linear vehicle model and subsequently for the nonlinear model where large lateral accelerations and disturbances are introduced. The simple model to be employed in the analysis is presented in Fig. 1.9. The track width has been neglected with respect to the radius of the cornering motion TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 17 which allows the use of a two-wheel vehicle model. The steer and slip angles will be restricted to relatively small values. Then, the variation of the geometry may be regarded to remain linear, that is: cosa - 1 and sina =a and similarly for the steer angle ~. Moreover, the driving force required to keep the speed constant is assumed to remain small with respect to the lateral tyre force. Considering combined slip curves like those shown in Fig.l.2 (right), we may draw the conclusion that the influence of Fx on Fy may be neglected in that case. In principle, a model as shown in Fig. 1.9 lacks body roll and load transfer. Therefore, the theory is actually limited to cases where the roll moment remains small, that is at low friction between tyre and road or a low centre of gravity relative to the track width. This restriction may be overcome by using the effective axle characteristics in which the effects of body roll and load transfer have been included while still adhering to the simple (rigid) two-wheel vehicle model. As has been mentioned before, this is only permissible when the frequency of the imposed steer angle variations remains small with respect to the roll natural frequency. Similarly, as has been demonstrated in the preceding section, effects of other factors like compliance in the steering system and suspension mounts may be accounted for. The speexl of travel is considered to be constant. However, the theory may approximately hold also for quasi-steady-state situations for instance at moderate braking or driving. The influence of the fore-and-aft force Fx on the tyre or axle cornering force vs slip angle characteristic (Fy, a) may then be regarded (cf. Fig. 1.9). The forces Fy~ and Fxt and the moment Mzl are defined to act upon the single front wheel and similarly we define Fy2 etc. for the rear wheel. In this section, the differential equations for the three degree of freeAom vehicle model of Fig. 1.4 will be derived. In first instance, the fore and aft motion will also be left free to vary. The resulting set of equations of motion may be of interest for the reader to further study the vehicle's dynamic response at somewhat higher frequencies where the roll dynamics of the vehicle body may become of importance. From these equations, the equations for the simple twodegree of freedom model of Fig. 1.9 used in the subsequent section can be easily assessed. In Subsection 1.3.6 the equations for the car with trailer will be established. The possible instability of the motion will be studied. We will employ Lagrange' s equations to derive the equations of motion. For a system with n degrees of freedom n (generalised) coordinates qi are selected which are sufficient to completely describe the motion while possible kinematic 18 TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY constraints remain satisfied",
+            " For more details we refer to Chapter 9 that is dedicated to short wavelength force and moment response. TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 23 From the equations (1.34b and c) the reduced set of equations for the two-degree of freedom model can be derived immediately. The roll angle q and its derivative are set equal to zero and furthermore, we will assume the forward speed u (= V) to remain constant and neglect the influence of the lateral component of the longitudinal forces Fxi. The equations of motion of the simple model of Fig. 1.9 for v and r now read: m((~ + u r ) = F 1 + F 2 (1.42a) I t ~- aFy 1 - b F 2 (1.42b) with v denoting the lateral velocity of the centre of gravity and r the yaw velocity. The symbol m stands for the vehicle mass and I (=Iz) denotes the moment of inertia about the vertical axis through the centre of gravity. For the matter of simplicity, the rearward shifts of the points of application of the forces Fy~ and Fy2 over a length equal to the pneumatic trail t~ and t2 respectively (that is the aligning torques) have been disregarded"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003413_978-4-431-53856-1_3-Figure3.2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003413_978-4-431-53856-1_3-Figure3.2-1.png",
+        "caption": "Fig. 3.2 Image of principles of Quad-Rotor MAVs flight. (a) Principle of roll control, (b) Principle of pitch control, (c) Motor rotation direction, (d) Principle of yaw control",
+        "texts": [
+            " Using three Gyro sensors and two micro controllers that are installed in the main Parameter Value body, it offers very steady flight even when it is controlled manually. The vehicle design consists of a carbon fiber airframe, so it has a light and strong body. The lift results from four brushless motor, and by using a special driver, the maximum speed reaches about 8,000 rpm. The complete vehicle is shown in Fig. 3.1. And the specification of X-3D-BL is shown in Table 3.1. The following explains the flight principles of a four rotor type MAV. The motors in Fig. 3.2a are numbered as No.1 at the front motor and then clockwise to No. 4 (also shown in Fig. 3.2c). Lift is obtained from the total force by all motors. The moment around X axis is generated by the rotational speed difference between No. 2 and No. 4 motors (Fig. 3.2a), so as the attitude angle around X axis of the airframe changes, the thrust is converted into the component force on Y direction. Using the same principle, by using the rotational speed difference between No. 1 and No. 3 motors, it is possible to control the X direction of the airframe (Fig. 3.2b). Moreover, the No. 1 and No. 3 rotors are rotating clockwise while the No. 2 and No. 4 rotors are rotating in the opposite direction (counter-clockwise) (Fig. 3.2c). The rotation around Z axis (Yaw) of the airframe is controlled by counterbalancing the moment (Fig. 3.2d). The internal PID controller of the airframe does the stabilization control by using three gyro sensors. In the manual control mode, the X-3D-BL behaves just like a radio controlled helicopter and it can be operated using a four channel Propo (RC controller). With attitude control mode, signals of the joystick are transmitted to the airframe as control instruction values of Altitude, Roll, Pitch, and Yaw angle. After the airframe receives the instructions, it calculates the appropriate response using its internal control system, and then sends the rotational speed instructions to the driver of each motor"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001305_s11274-007-9548-7-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001305_s11274-007-9548-7-Figure1-1.png",
+        "caption": "Fig. 1 Details of the apparatus used for enzyme immobilization (Adriano et al. 2005)",
+        "texts": [
+            " The support used for enzyme immobilization was activated carbon, SRD/21/1 grade, Mesh 10*35 US, from Speakman Carbons LTD. Polyethylene glycol (PEG 6000) was from Vetec and peptone (Bacto Pepetone with molecular weight range from 250 Da to 10 kDa) was purchased from Becton Dickson and Co. Other chemicals were of analytical grade. Preparation of immobilized enzyme Enzyme immobilization was obtained by contacting 3.0 mL of enzyme solution, in 25 mM phosphate buffer pH 6.0, with 0.1 g of activated carbon and stirred for 2 h 30 min at room temperature (28 C), using the apparatus presented in Fig. 1. Samples were taken along the time course of adsorption and hydrolytic activity on the bulk solution were measured. The adsorbed amount was calculated from the difference between enzyme hydrolytic activities before and after adsorption. After immobilization, the biocatalyst was separated by vacuum filtration and rinsed with 25 mM phosphate buffer pH 6.0 to remove any weakly adsorbed enzyme. Alternatively, the additives (PEG, BSA or peptone) were dissolved in phosphate buffer and added to the support simultaneously to enzyme immobilization"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003091_aim.2010.5695776-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003091_aim.2010.5695776-Figure3-1.png",
+        "caption": "Fig. 3. The phases of one step of the passive walking model with segmented feet. Each subfigure represents a walking phase during on step",
+        "texts": [
+            " If one of the forces decreases below zero in the direction orthogonal to the slope, the corresponding endpoint of the stance foot will lose contact with ground and the stance foot will rotate around the other endpoint. We choose parameters of the passive walker in normal range which lead to natural human-like gait. The motion will stop if the walker moves to a phase not included in this gait (For example, the heel of stance leg rises up before the swing foot contacts ground). One step can be described by several phases (see Fig. 3(a)-(h), each subfigure represents a walking phase during one step): 1) Phase a: The phase of foot rotation around toe joint. The rear foot rotates around toe joint while the toe keeps contact with ground. The toe will lose contact with ground when the ground force acted on the toe joint in the direction orthogonal to slope decreases to zero. Then the walker will move to phase b. 2) Phase b: The phase of foot rotation around toe tip. The toe rotates around the contact point. There is no constraint at toe joint in this phase"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001964_1.3300613-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001964_1.3300613-Figure1-1.png",
+        "caption": "Fig. 1. Diagram describing the loaded hoop on an incline. The hoop rolls in the x t ,y t frame and the mass P starts at =0 when released. Different / combinations were measured to test under what conditions the loaded hoop would hop.",
+        "texts": [
+            "4 In 2000 Theron did an extensive analysis of the hoop with a more realistic model in which the mass of the hoop is taken into account.5 He concluded the hoop would hop at an angle ranging from /2 to 2 . We created a quick demonstration of a rolling eccentrically loaded hoop to test the claims in Refs. 1\u20135. Because visual observations showed the hoop hopped at an angle clearly greater than /2, we conducted an experiment to test Theron\u2019s analysis. Theron modeled a perfectly rigid hoop with an attached point mass.5 Figure 1 shows the hoop with radius R rolling down a ramp, which makes an angle with the horizontal. Point P is the attached mass, is the angle that the mass has rotated from its starting position, and G represents the center of mass and is located a distance R from the geometric center, where is a measure of the eccentricity, = mattached mtotal . 1 Point O is the contact point between the hoop and the ramp where F and N act. F is the frictional force between the ramp and the hoop, and N is the normal force exerted by the ramp on the hoop"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000711_have.2008.4685294-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000711_have.2008.4685294-Figure2-1.png",
+        "caption": "Fig. 2. Voxel-based chip thickness evaluation.",
+        "texts": [
+            " Chip thickness is the most dominant factor in computation of the cutting forces and represented voxel-based in the next section. However, the chip thickness can be derived analytically and is affected by the feed rate of the tool [10]. The cutting force coefficients Kt and Kr are obtained from a set of orthogonal bone cutting tests [12],[13] and are equal to 230N/mm2 and 40N/mm2. Ka is a fraction of Kr to take into account the effect of the helix angle of the cutting edge and is equal to 20N/mm2. III. VOXELIZED FORCE MODEL A voxel-based force model is needed for the purpose of implementation. Fig. 2 shows the bone spherical tool in two successive positions - dashed and solid - where the spherical tool moves from the first position to the second position. As equation (1) is essentially discrete, only the chip thickness needs to be represented voxel-based to include all voxels between successive positions [10]. The chip thickness is determined proportional to the number of bone voxels located inside of the spherical tool. First, the voxels in the intersection region of the bone volume and the spherical tool are identified. Secondly, the Cartesian coordinates of each voxel v(X,Y,Z) are transformed into the cylindrical coordinates of the tool v(r,\u03a8, z) as shown in Fig. 2. Finally, the number of voxels are counted in each rotational angle increment and height increment, and then normalized based on the feed and assigned to the chip thickness. The simulation results showed a close force correlation between the voxel-based method and an analytical force model [10]. It is important to elaborate that because of the high spinning speed of the tool, each cutting edge travels a large angular distance at each 1msec servo period of the haptic loop. We have shown in our previous paper [10] that the total cutting force changes significantly as the angular location of the cutting edges change",
+            " Also, the bone volume extracted from CT or MRI data is a contiguous array of voxels. Figure (3) shows the flow chart of the simulation thread. First, the position of the center of the spherical tool (x0, y0, z0) is derived from forward kinematics of the haptic robot. Then, for each voxel of the tool (x1, y1, z1), collision is checked with the bone volume. The stiffness value of each voxel is set from CT data. The Cartesian coordinates of each voxel in the intersection region are then transformed into the cylindrical coordinates of the tool (as shown in Fig. 2). This cylindrical coordinate is derived from the forward kinematics of the haptic robot. We use a 2D array th[\u03a8, z] to store the chip thickness by adding up the voxel stiffness in each angle \u03a8 and height z. The force components (Fx, Fy, Fz) are calculated according to chip thickness matrix (th[\u03a8, z]) using equation (2). These forces are in the tool Cartesian coordinates and must be transformed to the global Cartesian coordinates for rendering. This volumetric method allows quantification of the mechanical properties of the interior structure of the bone"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-203-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-203-1.png",
+        "caption": "Figure 3-203: Typical arrangement of Sohre Turbomachinery, Inc. TYPE-A \u201cplunger type\u201d brush being used on a shaft end or collar. The brush can also be used radially, running on the shaft OD.",
+        "texts": [],
+        "surrounding_texts": [
+            "using a reliable gaussmeter with a hall probe, then the need for demagnetizing is clear. Following this, efforts should be made to maintain the equipment in a demagnetized state. Some of the suggestions that may be put into action are discussed below. 1. Existing Installations For the many units now installed and being manufactured, the prospect of residual magnetism causing problems is very real. Units should be maintained in a demagnetized state. Every effort should then be made to control and prevent remagnetization. \u2022 All components upon receipt following purchase, repair, or testing should be entirely free of residual magnetism. \u2022 Thorough deep-soaked demagnetization should be conducted on any component following magnetic particle inspection or on any component discovered to have high level of magnetic fields. Machinery Component Failure Analysis 291 \u2022 Welding on the compressor, turbine, or its piping should be controlled very carefully. The ground clamp and electrode cables should both be strung along the same path to the work area. Then the ground clamp should be connected to the same metal piece that is to be welded. \u2022 All components should have ground straps interconnected to the structure or station ground grid. The ground grid should have a ground resistance of less than 3 ohms. Also, lightning rods and other tall structures should have cables firmly interconnected to the ground grid. They further must be routed so they are not near nor do they link magnetic circuits such as closed-loop piping between the compressors or turbine to heat exchangers, condensors, boilers, etc. The goal is to provide a low-impedance discharge path for atmospheric discharge current but in such a way that component magnetization cannot occur. \u2022 Reliable brushes should be applied to the shafts to drain away electrostatic charge and to shunt persistent electromagnetic currents around bearing and seal surfaces. Brushes must be continuously conducting. This may require the use of a properly designed wire bristle brush (Figure 3-205). Close initial monitoring of the brush currents and voltages is necessary to assure that there is no compounding of magnetic fields due to the internal current paths. If this occurs, then brushes should be removed until components are demagnetized and there is assurance that remagnetizing currents are arrested. 2. Future Installations Remagnetization and occurrence of bearing damage is expected to continue unless significant measures are taken to correct the problems. Some of these measures are: \u2022 Insulation of all bearings, seals, couplings, and other components through which damaging currents now flow. \u2022 Installation of permanent brushes to shunt currents around the affected components. \u2022 Selection of materials for the equipment that is magnetically soft rather than magnetically hard. \u2022 Installation of coils in units at the time of original manufacture that can be used to effect demagnetization without having to disassemble the equipment."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000045_s11005-006-0069-3-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000045_s11005-006-0069-3-Figure1-1.png",
+        "caption": "Figure 1. a Complete integrability. b Kharlamova case.",
+        "texts": [
+            " In the case of integrability, due to the Liouville theorem, one can (locally) find trajectories (q\u00b1(\u03c4 ), p\u00b1(\u03c4 )) by quadratures (e.g., see [1]). On the other hand, the topological part of the theorem cannot be directly applied. Namely, if a compact connected component Lc of the regular invariant set Mc={(q, p)\u2208M | fi (q, p)= ci , i =1, . . . ,n\u22121} (18) does not intersect , then Lc is an (n\u22121)-dimensional torus with a uniform quasiperiodic dynamics in \u03c4 and therefore with a nonuniform quasi-periodic motion in the original time t . However, if Lc intersect , it may have a quite complicate topology (as an illustration, see Figure 1a). In the three-dimensional case, the topological structure of invariant manifolds of several integrable variants of the Suslov problem was studied by Tatarinov [24] and Okuneva [19,20]. Reconstruction. To reconstruct the motion (g(t), g\u0307(t)) on the whole phase space D, we have to solve the Poisson equations (4) for e1, . . . , en\u22121, i.e., to find all trajectories in D that projects to the given trajectory (en(t),\u03c9(t))= (q(t), Ap(t)). If (en(t0),\u03c9(t0)) is an equilibrium point or if (en(t),\u03c9(t)) is a periodic orbit, then the invariant set \u03c0\u22121{(en(t),\u03c9(t))}\u2282 D is foliated with invariant tori of maximal dimension rank SO(n\u22121) or rank SO(n\u22121)+1, respectively (e",
+            " The trajectories of the system can be found by quadratures and can be expressed in terms of elliptic functions of time t [13]. A generic invariant set Tc is diffeomorphic to R n\u22121 and T \u2217c is a disjoint union of 2l copies of (n\u22121)-dimensional closed balls (the number l, 0 l n\u22121 depends on the choice of constants ci ). Therefore, connected components of Mc are spheres. Trajectories of the vector field X H\u2217 pass through . Therefore, apart from a finite number of equilibrium points and their asymptotic trajectories, all the trajectories of the Suslov problem in the original time t will be closed (Figure 1b). Thus we get THEOREM 4. In the Kharlamova case, the phase space M of the reduced problem is almost everywhere foliated with (n\u22121)-dimensional spheres. The distribution D is filled up with conditionally-periodic trajectories of maximal dimension [n/(n\u22121)]+1. Klebsh\u2013Tisserand case. For generic values of the constants Ci and Bi the topology of invariant manifolds is much more complicated. Consider the quadratic potential v(q)= 1 2 (B1q2 1+\u00b7 \u00b7 \u00b7+ Bnq2 n ) (multidimensional Klebsh\u2013Tisserand case [13]) with Bi > Bn , i=1, "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000001_1.2032993-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000001_1.2032993-Figure1-1.png",
+        "caption": "Fig. 1 Friction",
+        "texts": [
+            " The paper describes the development of an eccentric planetary friction drive. The design configuration along with an improved wedge loading mechanism is presented. Finite element analysis FEA is performed to quantify design variables. Performance testing and evaluation are outlined to illustrate mechanical effi- ciency, torque capacity, and temperature characteristics. OCTOBER 2005, Vol. 127 / 85705 by ASME data/journals/jotre9/28735/ on 02/26/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Downloaded F Figure 1 shows the assembly of the current friction drive design. It is comprised of an outer ring, a sun roller, a loading planet, two supporting planets, and a carrier housing. The outer ring contains an inner cylindrical raceway and is connected through a spoke web to an output shaft that is supported by a double row bearing on the carrier housing. The sun roller contains an outer cylindrical raceway and is integrated with an input shaft that is supported by a ball bearing on the carrier. The sun roller is set eccentric to the outer ring by amount e, known as the eccentricity. Thus a wedge gap is formed in the annular space between the outer ring and the sun roller. For the orientation shown in Fig. 1, the wedge gap assumes a large end at top and a small end at bottom, as illustrated in Fig. 2. The loading planet is assembled in the wedge at the top position with its outer surface in frictional contact with the inner cylindrical raceway of the outer ring and the outer cylindrical raceway of the sun roller. The loading planet is \u201celastically\u201d supported through a bearing on a pin shaft that is fixed to the carrier. An elastic insert or a clock spring is provided between the pin shaft 858 / Vol"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000021_esda2006-95565-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000021_esda2006-95565-Figure10-1.png",
+        "caption": "FIGURE 10. CIRCULAR AND CUP BRUSHES FOR SURFACE FINISHING OPERATIONS",
+        "texts": [
+            " Nonetheless, and in contrast to the conclusions drawn by Peel [15], after an analysis of the information that has to do with vehicle velocity, it appears that flicking brushes could be more efficient than the cutting ones in practice, in terms of power consumption, as gutter brushes do not work stationary but travel at the sweeper speed. Therefore, the orientation angle of 30\u00b0 to 40\u00b0 is not necessarily the optimum. Surface finishing operations such as deburring and polishing are carried out for aesthetics or functional purposes, e.g., to eliminate burrs and sharp corners and improve surface finish [3]. These and other secondary material removal operations may be carried out by means of wire brushes [2] or abrasive (e.g., silicon carbide or aluminium oxide) -loaded nylon fibre brushes [11]. Figure 10 illustrates a circular brush and a cup brush, along with some brushing parameters. Although most of the research into the dynamics of brushing tools focuses on brushes for surface finishing operations, the published literature on these applications is still very restricted. Most of the research has been carried out on wire brushes (Fig. 10) by Stango and his co-workers [2-4,7]. Further, Fitzpatrick and Paul [11] deal with short-nylon fibreabrasive circular brushes. These works are reviewed below. Fitzpatrick and Paul [11] present theoretical and experimental results for the stiffness characteristics of shortnylon-fibre-abrasive brushes. The theory predicts that brush stiffness increases with brush penetration, which is the opposite of the findings by Stango and his co-workers (described below). As the equation is not derived in the paper, it is difficult to ascertain the reasons for this",
+            "asmedigitalcollection.asme.org/ on 02/03/2016 Te corresponding reduction of the elastic modulus, which in turn are caused by the higher temperatures developed. There are similar research studies that deal with circular wire brushes [2-4,10]. Stango et al. [2] derive equations, by using large displacement mechanics, to evaluate the forces, deformations, and stiffness of circular brushes with wire bristles of circular cross section that act against a flat, rigid, and frictionless workpart surface (Fig. 10.a, b, and d). In the model, the interaction between bristle and workpart is quasi-static, i.e., the bristle inertia forces, including centrifugal forces, are neglected; then, the higher the rotational speed of the brush, the lower the accuracy of the model. Other assumptions are: there is no eccentricity of the brush, no interaction among bristles, all the bristles are located radially, uniformly distributed along the circumferential direction, of equal length, and initially straight, and they withstand no axial deformation (i.e., along the bristle length). Shia et al. [3] and Heinrich et al. [4] studies are similar, but, contrary to Stango et al. [2], the former includes the effect of friction between bristles and workpart (Coulomb friction is assumed) and the latter analyses the effect of a workpart of constant curvature (convex surfaces of two different curvature radii and a flat surface are studied, as illustrated in Fig. 10.a). In addition to these three theoretical works, Stango et al. [10] study by means of experimental tests the characteristics of two brushes of the same type for three 6 Copyright \u00a9 2006 by ASME rms of Use: http://www.asme.org/about-asme/terms-of-use Do rotational speeds and a number of penetrations, and the results are compared with those of the model of Shia et al. [3]. Some results of the research works referred to in the previous paragraph are summarised as follows. The normal force between the workpart surface and the bristle tip is proportional to the flexural rigidity, EI, whereas bristle deformations are independent of it (provided Hooke\u2019s law is obeyed)",
+            " The reasons for this behaviour are not discussed; however, it may be due to the presence of stick-slip cycles at low rotational speeds and to phenomena related to bristle dynamics, as discussed in Section 3.2. It is noted that the COF exhibits some \u201coscillations\u201d (Fig. 12), which may be due to the induced bristle vibrations caused by their abrupt release from the workpart, because the vibration patterns and, in particular, the specific characteristics of the bristle dynamics when contact starts (point A in Fig. 10.a) depend on the rotational speed. Lastly, regarding the variation of the contact forces as the bristle travels along the surface, it is shown that they increase monotonically from zero (point A), up to a point called \u201crelease point\u201d (point B), where it is suggested that the bristle loses contact with the workpart and, thus, the forces disappear instantaneously. On the contrary, from about this point they do not vanish but fast decrease up to zero, and clearly the contact is lost only when the forces become zero",
+            " Further, due to the abrupt nature of the release of the bristles, the inertia forces, the induced bristle oscillations, and the propagation of these wnloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/03/2016 Ter oscillations between bristles may affect brush behaviour, but this is not studied in the works. Apart from the works on circular brushes, work has also been carried out on cup brushes. Through the model described in Section 2.3, Stango and Shia [7] analyse the characteristics of the bristles of freely rotating, straight (\u03c6 = 0 in Fig. 10.c) cup brushes. In the model, the bristles are of circular cross section (Fig. 10.d), and it is assumed that there is no eccentricity of the brush, there is no interaction among bristles, all the bristles are 7 Copyright \u00a9 2006 by ASME ms of Use: http://www.asme.org/about-asme/terms-of-use Dow uniformly distributed, of equal length, and initially straight, and that they withstand no deformation along the bristle length. The results indicate that, as expected, the maximum stress (at the clamped end) and the filament tip angle and displacement, as well as the operating brush diameter, increase with rotational speed and bristle eccentricity (mount radius); Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000496_ls.48-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000496_ls.48-Figure1-1.png",
+        "caption": "Figure 1. Hole-entry hybrid journal bearing (symmetric confi guration).",
+        "texts": [
+            " The present work was aimed to study the performance of hole-entry hybrid journal bearings by considering the variation of viscosity due to non-Newtonian behaviour of the lubricant and temperature rise of the lubricant. The analysis presented in the following subsection uses fi nite element method to model hole-entry journal bearing. The mathematical model involved simultaneous solution of Reynolds and energy equations in fl ow domain and heat conduction equations in solid domain (journal and bearing bush). The viscosity of non-Newtonian lubricant was assumed to follow the cubic shear stress law (\u03c4\u0304 + K\u0304 \u03c4\u03043 = \u03b3.\u0304). The results presented in the study are expected to be quite useful to the bearing designers. Figure 1 shows a geometric confi guration of a constant fl ow valve compensated non-recessed holeentry hybrid journal bearing. The journal is assumed to rotate with uniform angular velocity \u2126 about its equilibrium position. The Reynolds equation governing the fl ow of an incompressible lubricant in the clearance space of a journal bearing is defi ned in non-dimensional form as: \u2202 \u2202 \u2202 \u2202     + \u2202 \u2202 \u2202 \u2202     = \u2202 \u2202 \u2212           \u03b1 \u03b1 \u03b2 \u03b2 \u03b1 h F p h F p F F h o 3 2 3 2 11\u2126    + \u2202 \u2202 h t (1) where the non-dimensional viscosity function F\u03040, F\u03041 and F\u03042 are defi ned as: F d z F z d z F z z F F d zo0 0 1 1 0 1 2 1 0 1 1= = = \u2212{ }\u222b \u222b \u222b( / ) , ( / ) ( / ) ( / )\u00b5 \u00b5 \u00b5and Copyright \u00a9 2007 John Wiley & Sons, Ltd"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000456_j.triboint.2007.09.007-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000456_j.triboint.2007.09.007-Figure1-1.png",
+        "caption": "Fig. 1. Schematic illustration of flexible rotor supported by two couple-stress fluid film journal bearings.",
+        "texts": [
+            " The analysis focuses specifically on the effects of the lubricant type (i.e. Newtonian or non-Newtonian), the bearing suspension type (i.e. linear or non-linear), the rotational speed of the rotor and the degree of unbalance of the rotor. For simplicity, the analysis is based upon a shortbearing approximation. The dynamic equations of the rotor center and journal center are solved using the Runge\u2013Kutta method and the system response is illustrated by reference to bifurcation diagrams, dynamic trajectory diagrams and Poincare\u0301 maps. Fig. 1 presents a schematic illustration of a flexible rotor mounted in two pure squeeze couple-stress fluid film journal bearings with parallel, non-linear suspension systems. In this figure, Om is the center of gravity of the rotor; O1, O2 and O3 are the geometric centers of the bearing, the rotor and the journal, respectively; 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 are the stiffness coefficients of the springs supporting the two bearing housings; C1 is the damping coefficient of the supported structure; C2 is the viscous damping of the rotor disk; r is the mass eccentricity of the rotor; f is the rotational angle; R is the inner radius of the bearing housing and r is the radius of the shaft"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003174_robot.2010.5509512-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003174_robot.2010.5509512-Figure3-1.png",
+        "caption": "Fig. 3. Schema of simplified Gantry-Tau with points of TCP measurement set 1 (see Sect. III-B) and corresponding laser tracker positioning. Here, only every second measurement point layer is shown for better visibility (first layer to the left in the figure).",
+        "texts": [
+            " Their placement on plate and carts according to the so-called Tau configuration is such that the links belonging to one cluster form parallelograms, which assures a constant end- 978-1-4244-5040-4/10/$26.00 \u00a92010 IEEE 3709 effector orientation. The Gantry-Tau robot has thus three purely translational DOFs. A detailed solution of the kinematics problem can be found in [2] or [6]. Provided that the end-effector orientation is constant due to a perfect spherical joint placement, it is sufficient to consider the simplified robot shown in Fig. 3. The closure equation for link i is then (notation see Fig. 2): L2 i \u2212\u2016 T \u2212 sCi \u2016 2 2 = 0 (1) The cart position sCi of the simplified model can be expressed as: sCi = sC0 i + qi \u00b7 vi (2) where vi is the unit vector in positive track i direction. The track offset sC0 i of the simplified model is sC0 i = C0 i \u2212 RT \u00b7Pi (3) where Pi is link i\u2019s spherical joint position on the end-effector plate expressed in TCP coordinates and RT the rotation matrix between the TCP and the global frame. The nominal kinematic model assumes perfectly linear actuators and constant end-effector orientation guaranteed by the Tau-configuration of the spherical joints",
+            " In the following section, the assumptions of the nominal kinematics will be evaluated using the above measurements. Fig. 4 shows the end-effector orientation represented as ZYZ Euler angles along the grid of measurement set 1. The maximal Euler angle variations lie between 0.1\u25e6 (\u03b1) and 0.5\u25e6 (\u03b2 ). The repeating pattern exhibited can be associated with the 6 grid layers orthogonal to the actuator axes that the TCP is traversing (the robot is moving forward in one layer and the same path backwards in the next layer, Fig. 3). The pattern in Fig. 4 indicates that the TCP orientation errors are mainly caused by the arm system. Orientation variations of the carts along the guideways would have given variations between the layers for measurement points with the same Y -Z values (compare Fig. 3). To examine the actuator linearity and positioning accuracy, measurement set 2 was evalutated. Fig. 5 shows the absolute value of the residuals when fitting a linear function according to Eq. (2) between commanded cart positions qi(k) and measurements Ci(k) for carts 1 to 3. It can be seen that the linearity varies among the carts. The largest deviations are obtained for cart 2 (between 25 and 140 \u00b5m). In order to make an accurate model for these variations, measurements with a higher resolution along the guideway are necessary"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002929_s1068798x11010278-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002929_s1068798x11010278-Figure3-1.png",
+        "caption": "Fig. 3.",
+        "texts": [
+            " The reduced center of gravity O of the vibrational machine with the charged container reaches the orbit in Eq. (21) at t = ti and subsequently moves only over this trajectory. INTERACTION OF THE MOTION OF THE CONTAINER AND THE ABRASIVE MASS The total mass m is written in the form (22) where M0 is the mass of the abrasive granules; M1 is the mass of the moving parts of the vibrational machine (with the unbalanced mass). For the two masses, the positions of the centers of gravity are calculated on the basis of Fig. 3. A\u03c8 A2 2 B2 2+ ; \u03b4\u03c8 B2 A2 ,arctan= = A2 \u03b16' R1 \u03b15' R2\u2013( )/\u03940; B2 \u03b16''R2 \u03b12''R1+( )/\u03940.= =\u23a9 \u23aa \u23a8 \u23aa \u23a7 x t( ) Ax \u03c9t \u03b4x+( );sin= y t( ) Ay \u03c9t \u03b4y+( );sin= \u03c8 t( ) A\u03c8 \u03c9t \u03b4\u03c8+( )sin ,=\u23a9 \u23aa \u23a8 \u23aa \u23a7 x2 Ax 2 2xy AxBy \u03b4x \u03b4y\u2013( )cos\u2013 y2 Ay 2 + \u03b4x \u03b4y\u2013( )2 .sin= x2 Ax 2 y2 Ay 2 \u03c0 2 2k\u03c0,+ Ax Ay y Ax Ay y m M0 M1,+= C At time t0 = 0, the centers of gravity of masses M0 and M1 are concentrated at and , respectively. Then the total center of gravity of mass m is on the seg ment connecting points and . The position of the combined center of gravity (point O) at any time is given by Eq. (20). If we place the origin O of mobile coordinate system \u03beO\u03b7 at the center of mass m and the O\u03be runs along the container\u2019s symmetry axis, the seg ment will pass through the origin of coordinate system \u03beO\u03b7 at any time ti (Fig. 3). Using the equal torque theorem, we write (23) Hence (24) where and are coordinates of the points and in coordinate system \u03beO\u03b7. C0 0 C0 1 C0 0 C0 1 Ci 0 Ci 1 M0OCi M1OCi.= 0 0 M0 \u03bei M1 \u03bei ; M0 \u03b7i M1 \u03b7i ,= = 0 1 0 1 \u03bei, 0 \u03b7i 0 \u03bei, 1 \u03b7i 1 Ci 0 Ci 1 20 RUSSIAN ENGINEERING RESEARCH Vol. 31 No. 1 2011 F. S. YUNUSOV, R. F. YUNUSOV However, we may obtain Eqs. (23) and (24) on the basis of the equations of the centers of gravity in the mobile coordinate system (25) where and are the coordinates of points and in the system xO0y"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000016_6.2005-6088-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000016_6.2005-6088-Figure4-1.png",
+        "caption": "Figure 4. Virtual Structure Model",
+        "texts": [
+            " Control Method for Attitude Maneuver A. Virtual Structure A Virtual structure approach is one of approaches for conventional multi-spacecraft formation control16-19. In the approach, the entire formation is treated as a single structure. The control is derived in three steps in this approach. First, the desired dynamics of the virtual structure is defined. Second, the motion of the virtual structure is translated into the desired motion for each spacecraft, and finally, tracking control for each spacecraft is derived. Figure 4 is suggested to be a virtual structure model for the system as shown in Fig.3. The virtual structure consists of three rigid bodies that are placed on a concentric circle at 120 degrees intervals. The center of mass of the virtual structure and that of each rigid body are assumed to be connected by massless rods. The position of the C.M. of rigid body j measured in the coordinate frame V as follows. T jjj T j T j rrrr ][}{}{ 321vvr == (15) A scaling variable j\u03bb is introduced to represent expansion/reduction of the virtual structure"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002816_acc.2010.5530609-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002816_acc.2010.5530609-Figure7-1.png",
+        "caption": "Fig. 7. Tiger 72\u201d UAV platform.",
+        "texts": [
+            " The UAV could be first sent to fly at the flat area first to get an estimation of the prevailing wind and then be deployed to complex terrains for the further wind model estimation (future works). The real wind estimation data collected during the flight of a single UAV are used in this section to test the nominal wind estimation algorithm. 1) Tiger UAV Platform Introduction: AggieAir-Tiger UAV is used as the experimental platform for the wind data collection [21]. It is a flying-wing UAV with 72\u201d wing spans, as shown in Fig. 7. The weight of ChangE UAV is about 8 pounds including batteries, GPS, modem, autopilot, motors and sensors. Open source Paparazzi TWOG autopilot board is used as the autopilot hardware [22] with Microstrain Gx2 IMU and u-blox 5 GPS unit for inertial measurements. The specifications of AggieAir-Tiger UAV are described in Table I. All the sensor readings are sent to the ground control station (GCS) through a 900MHz serial modem in real time. The wind is estimated on the ground based on the UAV trajectory, the GPS speed, and the throttle percentage [22]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002483_kem.490.97-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002483_kem.490.97-Figure2-1.png",
+        "caption": "Figure 2. Schematic and photograph of Prototype I with top foil removed (see text for symbol definitions).",
+        "texts": [
+            " The measured frictional torque is a good indicator of the type of the developed lubricating film (full film or mixed lubrication) and allows determination of operating conditions at which full hydrodynamic film is developed. Since the test head (2) is supported by a hydrostatic spherical bearing, the frictional torque in the tested bearing can be precisely measured, even under substantial applied radial loads, using a calibrated elastic beam (5) and a non-contact displacement probe. Prototype I. Schematic view and a photograph of the first designed and built prototype is presented in Figure 2. In this prototype the elastic support consisted of 36 bump foil segments (3) positioned in 9 rows of 4. Each bump foil segment had 8 bumps (with the geometry defined in Table 1) and was attached at its left end (2) to the rigid steel half-sleeve support (1). Since water-lubricated bearings operate at low temperatures, gluing with high-strength industrial glue was attempted. The gluing operation was performed in a specially designed holder, which ensured proper positioning and firm contact between the glued parts",
+            " The first modification was the new method of attaching bump foils (3) to the half-sleeve (1), which eliminated the need of gluing. In the new design the bump foil sections (3) were mechanically attached to steel blocks (2), which in turn were placed in axial groves machined in the steel half-sleeve (1). The bump foil geometry remained unchanged (Table 1), which allowed the use of the existing hardware in the manufacturing process. However, due to the introduced modification, the locations of bump foil attachment changed for half of the bump foils (compare Figure 2 and Figure 3), resulting most likely in a significantly altered stiffness distribution along the angular direction [11]. The second important change, which was introduced, was replacement of the original steel top foil (coated with low-friction polymeric coating) with a thicker top foil made entirely of a commercially available low-friction polymeric foil (4). The way the top foil was attached to the half-sleeve (1) remained unchanged (5). The new top foil eliminated the danger of damaging the journal if the lubricating film of water is suddenly ruptured and a high-speed rub occurred"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000469_14644193jmbd78-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000469_14644193jmbd78-Figure4-1.png",
+        "caption": "Fig. 4 The elasto-multi-body dynamic model",
+        "texts": [
+            " In a multi-body model, the parts in the system are connected to each other by holonomic and nonholonomic constraints (various joints). These are listed in Table 2 for the single cylinder engine model. The total number of degrees of freedom (DOF) of the multi-body model is obtained using the Gurebler\u2013 Kutzbach expression as Number of DOF = flexible body modes(flywheel + crank + conrod) + 6 \u00d7 (number of rigid parts \u2212 1) \u2212 \u2211 (Contraints) = (56 + 66 + 32) + 6(16 \u2212 1)\u2212108 = 154 + 90 \u2212 108 = 136 Thus, the model comprises 136 DOF, including 90 rigid body motions and 154 of structural modal behaviour (Fig. 4). These are represented by a differentialalgebraic set of equations as d dt ( \u2202L \u2202q\u0307j ) \u2212 \u2202L \u2202qj + \u2202D \u2202qj \u2212 Fqj + n\u2211 k=1 \u03bbk \u2202Ck \u2202qj = 0 (1) where {\u03bej}j=1\u21926 = {x, y, z, \u03c8 , \u03b8 , \u03d5}T for the rigid body DOF, and {\u03bej}j=6\u21926+m = {x, y, z, \u03c8 , \u03b8 , \u03d5, q}T for the flexible bodies (q represents the modal coordinates and JMBD78 \u00a9 IMechE 2007 Proc. IMechE Vol. 221 Part K: J. Multi-body Dynamics at NATIONAL UNIV SINGAPORE on June 28, 2015pik.sagepub.comDownloaded from m their total number), L = T \u2212 V is the Lagrangian, which is the difference between kinetic and potential energies, and D = (1/2)q\u0307TZq\u0307"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002696_bf00534859-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002696_bf00534859-Figure3-1.png",
+        "caption": "Fig . 3. C l a m p e d T r u n c a t e d Conica l She l l .",
+        "texts": [
+            " The flow law (5b) gives _ ~ _ , ~ v (34) % 2 y,q, 2 k ~3 [(v \u00a7 w')/,'~] \" The bending moment m,~ must satisfy the yield condition (5 a): .,~ = 1 -- ,,'~ = 1 -- [ !'~ ~ (35) [2 , . # ' because of its continuity. For any other type of hinge circles, llv and z]w' may also be calculated from the velocity fields and conditions similar to (34) and (35) may be obtained explicitly. 6. Conical Shell. The present theory is also applicable to shell problems in other coordinates as, for instance, a cone (Fig. 3). Ins tead of F, the coordinate y = Y /L is used, but for easy reference ~v will be retained as a subscript. The equilibrium conditions for the cone are the following: 1 ~q,---v ' , s o - - - y (v - w t a n a ) , ~ = - k w'\" , ~o - (37) y The procedure of obtaining the general solutions is the same as before. Those for H+o and H~o are derived here for later use in an example. As in the general case, the flow laws (7b) and (8b) yield the relations (12). F rom (12b) and (37) one finds w\"/w\" =- - - 1/y whence w = B l l n y ~- B 2",
+            " - - m 0 = 4 - [ 1 - - 4 ~ k 2 ( B ~ - - y t a n c ~ ) 9'] , (41) and from (36 a) and (40) by simple quadrature % ~ 1 - - Y If (42) is substituted into (40) one obtains n0 = :J: 21k [ ~ (1 ~ - l n y ) - - Z y t a n a ] - - f p ~ d y -t- B 4 . (43) Equation (36b) may then be integrated to yield q~. Since qe and (m~ -- m0) are now known, (36c) may also be integrated to yield m~. 7. Clamped Truncated Conical Shell Subjected to a Line Load along its Shorter Edge. Let y = 1 and y = b be the inner and outer boundaries, respectively (Fig. 3). A line load with the dimensionless intensity q is applied at the inner boundary, normal to the shell. The solution is facilitated when the shell is shallow, because then the stress trajectory may be expected to be in the vicinity of that for a circular plate with similar boundary conditions. Since the velocity field of the plate is represented by a surface of revolution with negative Gaussian curvature, its stress trajectory lies on the side mo -- m~ = 1 of a Tresca hexagon for m~ and too. Its four-dimensional counterpart is the hypersurface H~o"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003064_978-3-642-34336-0_18-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003064_978-3-642-34336-0_18-Figure3-1.png",
+        "caption": "Fig. 3. Modeling and triangulation of workpiece",
+        "texts": [
+            " the coefficient a\u03c7 is a time invariant value therefore a positioning of the cutter, whereas u\u03c7 is linearly time-dependent: \u03c7 (\u03b1,\u03c9) = a\u03c7 + b\u03c7\u22c5(\u03b1\u2013\u03b1m) + c\u03c7\u22c5(\u03b1\u2013\u03b1m)2 + d\u03c7\u22c5(\u03b1-\u03b1m)3 + e\u03c7\u22c5(\u03b1\u2013\u03b1m)4 + f\u03c7\u22c5(\u03b1\u2013\u03b1m)5 + g\u03c7\u22c5(\u03b1\u2013\u03b1m)6+ p\u03c7\u22c5(\u03c9\u2013\u03c9m) + q\u03c7\u22c5(\u03c9\u2013\u03c9m)2 + r\u03c7\u22c5(\u03c9\u2013\u03c9m)3 + s\u03c7\u22c5(\u03c9\u2013\u03c9m)4+ t\u03c7\u22c5(\u03c9\u2013\u03c9m)5 + u\u03c7\u22c5(\u03c9\u2013\u03c9m)6 (1) In the simulation the workpiece and the tool envelope are modeled as 3D clouds of scattered points. With these points a mesh of triangles is generated for the workpiece and the tool. The modeling of the workpiece can be described in three steps, see figure 3. At first the cross section of the gear flank is defined by four points. With these points the gear width b, the toe and the heel of the bevel gear are defined. The heel is the face with the largest diameter respectively the largest distance to the central axis. The toe represents the face with the smallest diameter. In order to get a solid 3D body the cross section is revolved around the central axis by 360\u00b0. Finally this extruded body is getting triangulated as described by [3]. The tool envelope is modeled by an extrusion of the tool profile depending on the process kinematics, see figure 4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002824_2012-01-0908-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002824_2012-01-0908-Figure1-1.png",
+        "caption": "Figure 1. Outline of the PISDYN analysis model",
+        "texts": [
+            " The aims of this study, therefore, are to extract quantitative parameters related to oil film behaviors that correlate with pin noise by analyzing and simulating piston behaviors and, based on which, optimize the pin hole shape. In order to extract, from various design specifications of key dynamic systems, factors that affect piston pin behaviors, a piston behavior analysis tool was used to simulate pin behaviors. Ricardo PISDYN[3] was used for calculations. Calculations were conducted using the specs of the 1.5-liter engine with a semi-floating system used in the earlier study[1]. Figure 1 shows the outline of the model used in the analysis. For the piston and cylinder block, FEM was created to take into consideration elastohydrodynamic lubrication, just as in an actual engine. Measured values of combustion pressure, internal crank case pressure, and oil temperature were applied under the engine operating condition of 1850 rpm when hot at no load, a typical condition where pin noise is generated. Cylinder bore deformation reflects calculated thermal distribution and tightening distortion under operating conditions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.17-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.17-1.png",
+        "caption": "Fig. 1.17. Handling diagram resulting from normalised tyre characteristics. Equilibrium points I, 1I and Ill (steady turns) of which only I is stable, arise for speed V= 50 km/h and steer angle & = 0.04 rad. From the different line types the manner in which the curves are obtained from the upper diagram may be retrieved.",
+        "texts": [
+            " These characteristics subtracted horizontally from each other produce the 'handling curve'. Considering the equalities (1.80) the ordinate may be replaced by ay]g. The resulting diagram with abscissa al-a2 is the non-linear version of the right-hand diagram of Fig. 1.10 (rotated 90 o anti-clockwise). The diagram may be completed by attaching the graph that shows for a series of speeds V the relationship between lateral acceleration (in g units) ay[g and the relative path curvature 1/R according to Eq.(1.55). Figure 1.17 shows the normalised axle characteristics and the completed handling diagram. The handling curve consists of a main branch and two side lobes. The different portions of the curves have been coded to indicate the corresponding parts of the original normalised axle characteristics they originate from. Near the origin the system may be approximated by a linear model. Consequently, the slope of the handling curve in the origin with respect to the vertical axis is equal to the understeer coefficient ~/. In contrast to the straight handling line of the linear system (Fig.l.10), the non-linear system shows a curved line. The slope changes along the curve which means that the degree of understeer changes with increasing lateral acceleration. The diagram of Fig. 1.17 shows that the vehicle considered changes from understeer to oversteer. We define: 38 TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY understeer if: oversteer i f: R R (1.82) The family of straight lines represents the relationship between acceleration and curvature at different levels of speed. The speed line belonging to V= 50km/h has been indicated (wheel base l= 3m). This line is shifted to the left over a distance equal to the steer angle 6 = 0.04rad and three points of intersection with the handling curve arise",
+            " This opens the possibility to determine the handling curve with the aid of simple experimental means, i.e. measuring the steering wheel input (reduced to equivalent road wheel steer angle by means of the steering ratio, which method automatically includes steering compliance effects) at various speeds running over the same circular path. Subtracting normalised characteristics may give rise to very differently rear). shaped handling curves only by slightly modifying the original characteristics. As Fig. 1.17 shows, apart from the main branch passing through the origin, isolated branches may occur. These are associated with at least one of the decaying ends of the pair of normalised tyre characteristics. In Fig. 1.18 a set of four possible combinations of axle characteristics have been depicted together with the resulting handling curves. This collection of characteristics shows that the nature of steering behaviour is entirely governed by the normalised axle characteristics and in particular their relative shape with respect to each other",
+            " Instead of the cornering stiffnesses C defined in the origin of the tyre cornering characteristics, the slope of the normalised characteristics at a given level of ay/g becomes now of importance. We define Q~). 10F i - ( i - 1 , 2 ) ( 1 . 8 3 ) ' F . Oa. Zl l The conditions for stability, that is: second and last coefficient of equation comparable with Eq.(1.47) must be positive, read after having introduced the radius of gyration k (k2= I/m): (k 2 +aZ)qs1 +(k 2 +bZ) q52 > 0 (1.84) ~1~2 ( 0 ~ ) >0 (1.85) Ol/R v The subscript V refers to the condition of differentiation with V kept constant, that is while staying on the speeA line of Fig. 1.17. The first condition (1.84) may be violated when we deal with tyre characteristics showing a peak in side force and a downwards sloping further part of the characteristic. The second condition corresponds to condition (1.65) for the linear model. Accordingly, instability is expected to occur beyond the point where the steer angle reaches a maximum while the speeA is kept constant. This, obviously, can only occur in the oversteer range of operation. In the handling diagram the stability boundary can be assessed by finding the tangent to the handling curve that runs parallel to the speeA line considered. In the upper diagram of Fig. 1.20 the stability boundary, that holds for the right part of the diagram (ay vs l/R), has been drawn for the system of Fig. 1.17 that changes from initial understeer to oversteer. In the middle diagram a number of shifted V-lines, each for a different steer angle fi, has been indicated. In each case the points of intersection represent possible steady-state solutions. The highest point represents an unstable solution as the corresponding point on the speeA line lies in the unstable area. When the steer angle is increased the two points of intersections move towards each other. It turns out that for this type of handling curve a range of fi values exists without intersections with the positive half of the curve",
+            "94) with slope d%/dal kept constant. The following three isoclines may already provide sufficient information to draw estimated courses of the trajectories. We have for k 2 = ab: vertical intercepts (da 2/da~ ~ oo): _ gl F l(al) a 2 + a 1 - d (1.95) V 2 F 1 horizontal intercepts (da2 /da~ ~ 0): _ gl Fz(a2) 0~ 1 - - + 0~ 2 + d ( 1 . 9 6 ) g 2 F 2 intercepts under 45 ~ (da2/dal = 1): f l ( O ~ l ) - F 2 ( o ~ 2) (1.97) F, F2 Figure 1.22 illustrates the way these isoclines are constructed. The system of Fig. 1.17 with k = a = b, ~ = 0.04 rad and V= 50 km/h has been considered. Note, that the normalised tyre characteristics appear in the left-hand diagram for the construction of the isoclines. The three points of intersection of the isoclines are the singular points. They correspond to the points I, II and III of Fig. 1.17. The stable point is a focus (spiral) point with a complex pair of solutions of the characteristic equation with a negative real part. The two unstable points are of the saddle type corresponding to a real pair of solutions, one of which is positive. The direction in which the motion follows the trajectories is still a question to be examined. Also for this purpose the alternative set of axes with r and v as coordinates (multiplied with a factor) has been introduced in the diagram after using the relations (1",
+            "24 the influence of an increase in steer angle ~ on the stability margin (distance between stable point and separatrix) has been shown TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 49 for the two vehicles considered in Fig. 1.20. The system of Fig. 1.23 is clearly much more sensitive. An increase in 6 (but also an increase in spee~ V) reduces the stability margin until it is totally vanished as soon as the two singular points merge (also the corresponding points I and II on the handling curve of Fig. 1.17) and the domain breaks open. As a result, all trajectories starting above the lower separatrix tend to leave the area. This can only be stopped by either quickly reducing the steer angle or enlarging 6 to around 0.2rad or more. The latter situation appears to be stable again (focus) as has been stated before. For the understeered vehicle of Fig. 1.24 stability is practically always ensured. For a further appreciation of the phase diagram it is of interest to determine the new initial state (ro, Vo) after the action of a lateral impulse to the vehicle (cf",
+            " Obviously, we find that at braking of the front wheels these components will counteract the cornering effect of the side forces and thus will make the car more understeer. The opposite occurs when these wheels are driven (more oversteer). For a more elaborate discussion on this item we may refer to Pacejka (1973b). At hard braking, possibly up to wheel lock, stability and steerability may deteriorate severely. This more complex situation will be discussed in Chapter 3 where more information on the behaviour of tyres at combined slip is given. Possible steady-state cornering conditions, stable or unstable, have been portrayed in the handling diagram of Fig. 1.17. In Fig. 1.22 motions tending to or departing from these steady-state conditions have been depicted. These motions are considered to occur after a sudden change in steer angle. The potential available to deviate from the steady turn depends on the margin of the front and rear side forces to increase in magnitude. For each point on the handling curve it is possible to assess the degree of manoeuvrability in terms of the moment that can be generated by the tyre side forces about the vehicle centre of gravity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000809_robot.2008.4543556-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000809_robot.2008.4543556-Figure2-1.png",
+        "caption": "Fig. 2. Illustration of path following (w/o sideslip)",
+        "texts": [
+            " 1, the slip angle of the vehicle, \u03b20, is calculated by using the longitudinal and lateral linear velocities on the vehicle coordinate, vx and vy , as follows: \u03b20 = tan\u22121(vy/vx) (1) Based on the kinematics and terramechanics, non-zero slip angle also occurs due to cornering effect coupled with 978-1-4244-1647-9/08/$25.00 \u00a92008 IEEE. 2295 wheel slippages. The wheel slippage can be divided into the longitudinal and lateral slips. A key issue in our proposed control is the compensation of these three slips while the rover follows an arbitrary path. A general illustration of the path following problem is shown in Fig. 2. In the path following problem, a feedback control law is employed to reduce both distance and orientation errors. The distance error, denoted by le, is determined as a distance between P and Pd. The orientation error, \u03b8e, is given as \u03b80 \u2212 \u03b8d. Here, \u03b80 is a vehicle orientation around yaw-axis of the vehicle. \u03b8d is the angle between the x-axis of the inertial coordinate system and the tangent to the path at Pd (vehicle\u2019s desired orientation). Referring to conventional path following algorithm, a vehicle is controlled such that the vector of linear velocity of the vehicle v0 coincides with a tangent of a given path as shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001483_09544070jauto781-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001483_09544070jauto781-Figure1-1.png",
+        "caption": "Fig. 1 Finite element model of a connecting-rod bearing Fig. 2 Finite element model of a main bearing",
+        "texts": [
+            " When dividing the model into elements, the bush is divided into hexahedral elements and the other parts of the connecting rod are divided into tetrahedral elements. The element division on the bush surface is controlled especially to make nodes on the bush surface correspond to nodes of the difference grid used to calculate the oil-film pressure of the bearings and to assure that the oil-film pressure of bearings can be applied correspondingly to the bush surface. The finite element model of the connecting-rod bearing is shown in Fig. 1, which consists of 21 775 elements and 36 723 nodes. The nodes on the transitional section between the small end of the connecting rod and the body of the connecting rod are assumed fixed. (b) Main bearing. The finite element model of the main bearing is shown in Fig. 2 and is composed of the bush, the main bearing cap, and the top half of the main bearing housing in the cylinder block. The bush is divided into hexahedral elements and the other parts of the main bearing housing are divided into tetrahedral elements"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003229_s11431-012-4986-3-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003229_s11431-012-4986-3-Figure2-1.png",
+        "caption": "Figure 2 Test gears: Type A gears.",
+        "texts": [
+            " The bulk temperature could be immediately measured by a handheld infrared thermometer through the top hole of the test gearbox when a test finished at each load stage. The rotation speed of the motor was 1450 rpm while the circumferential speed at the pitch point of the test gear was about 8.3 m/s. The test procedure was in line with GB/ T19936.1-2005, Section 1: A/8.3/90. To accelerate the occurrence of scuffing failure, the test gears were specially designed [9] with a small width to increase the local contact pressure and high relative speed. In this study, gears of type A (Figure 2.) were used as the test gears. The test oil was one of the potential oils, instead of the synthetic oil. The main properties of the oil are listed in Table 1. Four levels of oil immersion depth illustrated in Figure 3 were used in the test to find a reasonable lubricant quantity. Level-1 was at the center of the shaft which was the standard oil\u2019s level (or the reference level) in FZG test. Level-4 was equal to the height of one tooth of the gear which was usually considered as the minimum quantity of dip-lubricated gear transmission in practice"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001895_09544070jauto1066-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001895_09544070jauto1066-Figure9-1.png",
+        "caption": "Fig. 9 Schematic diagram of the installation of the displacement measuring system for the DMF on the engine",
+        "texts": [
+            " The error between the output of the CVDT telemetry system and the second-order polynomial fit value was within 0.86 per cent of the full scale. Consequently, the maximum margin of error of the CVDT telemetry system was within 0.87 per cent: the square root of the sum of squares between the nonlinearity error and the error stemmed from the temperature coefficient of resistance. In this study, a 2.0 l manual transmission vehicle is used. The internal combustion engine is a spark ignition engine and its geometric compression ratio is 10.5. Figure 9 shows the CVDT-mounted DMF between the engine and transmission. Two concentric coils were mounted on the DMF to supply the power. The receiver is located in the hollow space between the engine and the manual transmission. The receiver part, which receives the data from JAUTO1066 Proc. IMechE Vol. 223 Part D: J. Automobile Engineering at LAURENTIAN UNIV LIBRARY on November 25, 2014pid.sagepub.comDownloaded from the telemetry, is connected to a PC, and the angular displacement of the DMF is then measured"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001544_6.2008-6842-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001544_6.2008-6842-Figure2-1.png",
+        "caption": "Figure 2. Position view of the chase UAV and the target aircraft.",
+        "texts": [
+            " Although the proportional navigation (PN) guidance might be also considered as well, the PN guidance is prone to guide away from the target when the closing velocity is negative (the target velocity is greater than that of the UAV.) On the other hand, the pure pursuit guidance always guides the UAV to orient in the target direction in spite of both velocities. Therefore, we use the pure pursuit guidance in this study. Assume that a chase UAV with the velocity of CV pursues a target aircraft as shown in Fig. 2. The following equation is satisfied if the line of sight (LOS) vector R and the chase UAV's velocity vector CV are heading towards in the same direction. 0RVC =\u00d7 (19) In order to guide the chase UAV into the pure pursuit course, the following acceleration feedback is required so that Eq. (19) can hold. RV N C CC d VRV a \u00d7\u00d7 = )( (20) The following equation might be preferable instead of using Eq. (20) for implementation purpose in the real system, especially for large LOS angle in applying the real LOS angle feedback"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003763_s10068-012-0158-2-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003763_s10068-012-0158-2-Figure4-1.png",
+        "caption": "Fig. 4. Response surface and contour plots showing the effect of the content of enzyme and reaction temperature on the molar conversion yield at 15 of molar ratio of lauric acid to erythorbic acid.",
+        "texts": [
+            " Molar conversion was enhanced by increasing the molar ratio of lauric acid to erythorbic acid. At a fixed molar ratio of lauric acid to erythorbic acid, varying enzyme content had a little effect on the molar conversion yield. Figure 3 shows that a high molar conversion yield (over 60%) could be obtained using a high molar ratio of lauric acid to erythorbic acid (over 20) regardless of the amount of enzyme used. These data indicate that the substrate molar ratio was the most important variable. Figure 4 shows the effect of the enzyme content (1,000-5,000 PLU) and reaction temperature (25-65oC) on the synthesis of erythorbyl laurate at a molar ratio of lauric acid to erythorbic acid of 15. At a fixed enzyme content, the molar conversion yield increased rapidly when the temperature reached approximately 50oC, and it then leveled off. At a fixed temperature, the molar conversion yield was varied slightly with increasing enzyme content, especially when the temperature exceeded 50oC. This result indicates that the reaction temperature had the greatest effect on the molar conversion yield"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001879_j.jappmathmech.2010.01.001-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001879_j.jappmathmech.2010.01.001-Figure1-1.png",
+        "caption": "Fig. 1.",
+        "texts": [
+            " The analytical results obtained in this paper supplement the numerical investigations of this problem performed in Ref. 4. Consider a heavy rigid body on a horizontal plane. The body (which we will call a top) consists of two spherical segments with radii r1 and r2, which complement one another in the sense that while the first segment is formed by rotating a 2( \u2212 ) arc of a circle about the axis of symmetry, the other is formed by rotating a 2 arc. The two segments are rigidly joined by a rod that passes through their centres O1 and O2 (Fig. 1). The geometrical parameters of the top , r1, r2 and l (l is the distance between the centres of the segments O1 and O2) are related by the expression Prikl. Mat. Mekh. Vol. 73, No. 6, pp. 867-877, 2009. E-mail address: azobova@mail.ru (A.A. Zobova). 0021-8928/$ \u2013 see front matter \u00a9 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.jappmathmech.2010.01.001 To be specific, we will assume that r1 > r2 and that \u2208 (0, /2). The top is then supported on the horizontal plane at a point on the first spherical segment if \u2208 [0, \u2212 ) and at a point on the other spherical segment if \u2208 ( \u2212 , ], and the top is supported on the plane at two points if = \u2212 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003586_icinfa.2013.6720328-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003586_icinfa.2013.6720328-Figure8-1.png",
+        "caption": "Fig. 8. 6-DOF arc welding industrial robot",
+        "texts": [
+            " \u03c9 \u03b1 = \u2212 \u2212 \u0394accT dt T (18) Step4: If the total distance D is smaller than min 2D ( min 2\u2264D D ), this step will be executed. Because the distance is so short that the maximum of acceleration could not be achieved, the new acceleration and velocity will be recalculated. 2 max3 2 2 .\u03b1 \u03c0 = DJ .= =accinc accdecT T dt 0.= =vel accT T (19) A model of 6-DOF arc welding industrial robot has been used to test the different performances of orientation interpolations between Euler angles and unit quaternion proposed in this paper. 6-DOF arc welding robot shows in Fig.8 and Fig.9. MATLAB/SIMULINK is employed to build the simulation platform of industrial robot. The inverse kinematic, trajectory planning and inverse dynamic are parts of robot control system illustrated in Fig.10. The performances including torque, and joint velocity of robot were compared. The dynamic parameters of robot are stated in Table I. The rotation demand is given in three different forms as follows, including Euler angles, unit quaternion and rotation matrix. Euler angles: 30 , 60 , 90\u03c6 \u03b8 \u03d5= \u2212 \u00b0 = \u00b0 = \u00b0 Quaternion: [ ]0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001199_0278364907085565-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001199_0278364907085565-Figure8-1.png",
+        "caption": "Fig. 8. Modular manipulator.",
+        "texts": [
+            " Therefore, all of the terms qi k (successive to q1 1) also become very small scalars and, hence, by collecting a finite number of them, any risk of obtaining joint velocity references that are too high is certainly avoided. By taking into account the previously devised iterative algorithm for kinematic inversion, we can now shift our attention back towards the original problem of controlling the motion of self-reconfigurable modular manipulators through a totally decentralized control architecture. To this aim, the case of a. serial chain with n degrees of freedom composed of a set of atomic robotic modules with only one degree of freedom, as depicted in Figure 8, is considered here (without loss of generality) as a representative example. With reference to Figure 8, for every ith robotic unit an end-effector frame ei and a base frame bi may be defined (with an obvious meaning), whose relative position and attitude are described by the \u201cinternal\u201d transformation matrix bi ei T qi , depending on the sole joint variable qi , as well as on the local geometry of the considered module. In addition, all of the above frames satisfy the following coincidence relationships: b1 0 ei bi 1 en t i 1 2 n 1 (38) reflecting the way the modules are assembled together. The required motion task is defined in terms of making the overall tool frame t asymptotically converge towards an assigned goal frame g ",
+            " In addition, parallel experimental research activities have been oriented to develop two robotic prototypes a 3D modular manipulator with five degrees of freedom and a 2D tree-structured chain with nine degrees of freedom. Details on both the simulative results and the experimental activities are provided in this section. As concerns the tests performed on serial chains, the present work shows the results of three different simulations performed on the model of a 3D modular manipulator with eight degrees of freedom similar to the one depicted in Figure 8. As a preliminary task (see Extension 1), an end-effector point-to-point movement, from the initial position of [0.4, \u2013 0.6, 1.3] m with respect to the base of the robot towards the point [\u20130.4, \u20130.8, 0.7] m, while maintaining the initial endeffector orientation, has been executed. Different trials have at CARLETON UNIV on March 16, 2015ijr.sagepub.comDownloaded from been performed, by varying the number of iterations computed within an overall sampling interval T of 5 ms. In Figure 12 the 3D trajectories actually followed by the end-effector during the different trials are reported (solid lines) and compared with the ideal straight trajectory (dashed)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000413_0-387-23335-0_3-Figure3.18-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000413_0-387-23335-0_3-Figure3.18-1.png",
+        "caption": "Figure 3.18. Schematic illustration\" of the sensor head of a YSl 8500, wet optical sensor for carbon dioxide that is currently on the market.",
+        "texts": [],
+        "surrounding_texts": [
+            "This article began witli a brief look at the possible areas of application of optical sensors for carbon dioxide. A number of studies, carried out using dry and wet luminescent optical sensors for carbon dioxide, have illustrated the efficacy of these indicators for the detection and measurement of carbon dioxide in: blood\"'', food packages' , bioreactors'''\" and seawater'\". However, despite this success and promise, these sensors have been slow to take off as commercial products and the detection and analysis of carbon dioxide is still dominated by the Severinghaus electrode (dissolved work) and infrared spectroscopy (gaseous work). The reasons for this poor transition from research bench to market place are numerous but include: consumer resistance to new, and largely still unproven technology and, more seriously, basic concerns regarding the technology itself Thus, as we have seen intensity-based measurements are fraught with several niggling and undermining problems, such as signal drift and dye bleaching, although many can be eliminated using a wavelength ratiometric method\"'' as we have seen. Lifetime-based measurements were very expensive and are still considered so for many who might otherwise readily adopt this technology, such as research laboratories, despite the recent notable inroads made by workers using long-lived donor lumophores'\"'''^l All carbon dioxide optical sensors are also temperature sensitive, most markedly so, and many will exhibit some sensitivity to changes in humidity or osmotic pressure. Most optical sensors for dissolved carbon dioxide measurements require a gas-permeable membrane cover, to prevent dye-leaching and ion-exchange taking place, both of which can cause such sensors to fail. Despite the above technological concerns, there are one or two examples of apparent commercial success in the transfer of optical sensors for carbon dioxide technology from the laboratory to the market place. Thus,",
+            "156 MILLS \u00a37/1 \u00a3.\nYellow Springs Instrument (YSl) currently promote a wet, luminescence intensitybased carbon dioxide monitor that allows the precise, real-time measurement of dissolved carbon dioxide in situ'\". Their YSl-8500 instrument has a range of 1-25% COT and an accuracy of typically \u00b1 5% of reading. Their sensor system exhibits only a 2% drift per week but, as you might expect given its wet nature, has a long 90% response time of <7 minutes with the recovery even longer, due to the hyperbolic response characteristics of such sensors. The sensor is not very bulky (12mm diameter with a 70-320 mm insertion depth) and utiUses HPTS as the pH-sensitive lumophore. The YSl systems employs a ratiometric''' analysis of the dye's fluorescence, exciting at two different wavelengths and monitoring the luminescence intensity at one, in order to minimise the various instrumental effects associated with such intensity-based measurements. Fig. 18 provides a schematic illustration of the sensor head for the YSl-8500, highlighting its major features. The system uses replaceable, disposable sensor capsules, thus promoting one of the attractive features of most optical sensors; the disposable nature of the transducer element.\nWhereas YSl have opted for a wet, intensity-based optical sensor for carbon dioxide, OceanOptics, with their FC02-R fibre optical sensor, have opted for a dry sensor which encapsulates the pH-sensitive dye, HPTS, in a sol-gel medium, and is covered by a black silicone GPM coating'''. The sensor has a dynamic range of 0- 25% CO2 and a resolution of at least 0.03%. The instrumentation does not use a ratiometric technique (i.e. no two excitation sources) to interrogate the sensor film as employed by YSl, but instead monitors the whole luminescence spectral output using a miniature spectrophotometer. Like the YSl-8500, the FC02-R uses a blue LED as the excitation source. The response time of the FC02-R is typically 10 minutes and the probe itself is quite small (1.6 mm) diameter. However, as might be expected for a sol-gel based sensor, it must be stored in water at least 2 days before use and it is recommended to store it subsequently in water at all times after.\nIt is possibly surprising that there are no major lifetime-based optical sensors in the market, despite the fact that lifetime based optical sensors for oxygen already",
+            "exist (e.g. OxySense''\"\"). However, it is most likely that the higher cost of such teclinologically advanced systems is still proving a major barrier to its marlcet transition. The recent move towards cheaper, lower frequency, diode-based phase modulated systems, whether they be FRET- or DLR-based, may eventually lead to a relatively inexpensive, commercially viable product that challenges and widens the market for small intensity-based optical sensors for carbon dioxide. However, for this to happen a great deal more research is necessary. It remains to be seen, therefore, what the commercial future of optical sensors for carbon dioxide is which a slightly disappointing conclusion is given its history of initial rapid development and the importance of the area of analysis to life, industry and the environment."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001232_j.cmpb.2007.03.004-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001232_j.cmpb.2007.03.004-Figure9-1.png",
+        "caption": "Fig. 9 \u2013 A mass point module.",
+        "texts": [
+            " ore in detail, in [18,19] two distinct neural computational nits (called modules) are used, each computing the dynamics f a particular element of the physical model. The first modle (mass point module) is composed by neurons that compute he dynamics of the mass points, whereas the second modle (spring module) is composed by neurons that compute the ynamics of the springs. These modules are connected together according to the hysical model. The resulting neural network is recurrent, since here are self-connected neurons (called integrator units, see 18]). More in detail, the mass point module consists of three istinct neurons as shown in Fig. 9. These neurons compute espectively the acceleration, the velocity and the position of each ass point. On the other hand, the spring module (Fig. 10) computes he instantaneous reaction force engendered by a stressed pring. This force depends on the position and the velocity f the uttermost points of the spring. Note that, in Fig. 10, he weights of the connections to the neuron that computes he instantaneous reaction force (neuron F in Fig. 10) repreent the constants characterizing the spring: the elastic (k in ig",
+            " The simulation engine comprises the following classes: \u2022 The class simulation, that implements the simulation steps. \u25e6 [ \u25e6 bullet] It computes the internal forces using the class spring (see below). \u25e6 [ technique architecture. \u25e6 bullet] It applies the external forces. \u25e6 [ \u25e6 bullet] It computes the new positions and new velocities of the mass points using the class mass (see below). \u25e6 The class mass that implements the mass point module (see Section 3.2). The module, representing the mass point dynamics consists of three distinct neurons as shown in Fig. 9. These neurons compute respectively the a cceleration, the velocity and the position of each mass point (method simulate ( )). \u25e6 The class Spring that implements the spring module (see Section 3.2). The spring module (Fig. 10) computes the instantaneous reaction force engendered by a stressed spring (method solve ( )). The reaction force depends on the position and the velocity of the uttermost points of the spring. Fig. 11 shows the overall class architecture. We have omitted the detailed description of class variables and methods, since they are available at [20]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002285_660432-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002285_660432-Figure9-1.png",
+        "caption": "Fig. 9 - Force dynamometer details, constant torque machine",
+        "texts": [
+            " From these photographs it may be seen that the small platen which moves up and down in an oscillatory fashion, transmitting the unbalanced eccentric force, is coupled by means of a torque arm directly through a part similar to arm C of Fig. 1. This, in turn, is held to a flat plate by means of the test bolt. The flat plate is securely fixed to the base of the testing machine by means of two heavy supports, as shown in Fig. 7. Fig. 8 shows an element of the test fixture corresponding to part C of Fig. 1. This rod is, of course, threaded to receive the test bolt and is further specially constructed to measure the axial load in that bolt. The form of this construction is shown in Fig. 9, where it is seen that a portion of the cylindrical member is removed so that a thin cylindrical shell can be substituted. This shell has been instrumented with strain gages so that axial load may be accu- rately measured. TESTING PROCEDURE The procedure for testing a bolt under a particular set of conditions begins by inserting the bolt through the fixed mem - ber and into the threaded cylindrical portion, followed by tightening the bolt to some level of shank stress as indicated by the strain gage instrumented dynamometer on the shoulder of the cylindrical portion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000872_bf00191104-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000872_bf00191104-Figure4-1.png",
+        "caption": "Figure 4.",
+        "texts": [
+            "1) @~= a,, I - ( l_v ) [4 f l x f i z_ ( l +fl2)2] so that at a sub-Rayleigh velocity of wedge motion the stress a\u00a2~ is less than a,, and at a superRayleigh velocity o-\u00a2~ is higher (by absolute value) than a,,. Thus maximum tangential stress at the continuation of the cut appears to differ from zero: ~'max(~, 0) = a'tt(~' 0)--O-~(~, 0) 2 ,~ 0 . (4.2) It follows that since the front edge of the wedge has though a small but finite radius of curvature there arises a plastic zone ahead of the wedge (Fig. 4)*. Strictly speaking it implies the necessity of some modification of the statement of the problem: assuming that the plastic zone is thin (its thickness is of the order no higher than the wedge thickness) and taking again the boundary conditions at the symmetry line we obtain for a lower half-plane a modified boundary value problem aCn(~,0)=0 ( - ~ < ~ < ~ ) , u,(~,0)=-0 ( - ~ < ~ < 0 ) (4.3) Zmax(~, 0) = Zr (0 < ~ < I'), U,(\u00a2, 0) = --f(~) (t' < ~ < t), a,,(\u2022, 0) --= 0 (4 > I) where l' is the length of plastic zone, the coordinate origin is moved to the leading edge of the plastic zone (Fig. 4), zy is the yield stress (in compression) of the wedged body material, the stresses in the points \u00a2 = l' and ~ = l are to be finite. Due to (4.1) the condition Zmax(~, 0)= Zr(0< ~ < /').can be rewritten as 2(1 - v) [4fl1 f12 - (1 + flz)a] ann (4, 0) = m 2 (1 + ~2) - - r r (0 < ~ < l') (4.4) so that the problem (4.3) reduces to a mixed problem of analytic function theory for a half-plane. Obtaining the solution of this problem in closed form gives no difficulties; we shall not however present it here, suffice it to say that it would imply a plastic zone form with a sharp front edge due to the finite stress at the point \u00a2 = 0, q = 0. This means that the wedge is so-to-say built on a plastic \"beak\" (Fig. 4) and becomes sharp. Including into the wedge the plastic \"beak\" we can use the solution of the above problem and obtain the relation for the contact region length 4zr (1 - vZ) fll _ l~f I f'(~)d~ (4.5) E ( l + f l 2) 0 ~(1-~) ~ which follows from the condition of equality of maximum tangential stress at the end of the * In prinople due to high pressures and rates of deformation a zone of intensive heating can be supposed to arise ahead of the wedge favouring the appearance of a plastic flow and even of a melting zone"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003684_0016-0032(63)90563-5-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003684_0016-0032(63)90563-5-Figure7-1.png",
+        "caption": "FIG. 7. Nondimensional displacement versus x: theoretical for = 1, 2, 5 and experimental.",
+        "texts": [
+            " Ow Since v = aT' the displacement for r = r0 (constant) can be de- termined as /: w(r0, t) = v(ro, ~)d~. (17) ing the velocity from Eq. 9 into Eq. 17 and performing the integrat ion with respect to c0 we obtain ~(xo, z) = fo ~ Ji(7)Jo(TXo) (1 -- e-*,2)d7 ~[ 2 , ( i s ) where the reduced variables of Eq. 12 have been introduced together with = XR (19) w '~ - - (nondimensional displacement) . VoR~ ~_ P The above expression has been integrated numerical ly for r = 1, 2, 5 and plot ted as shown in Fig. 7. I t can be noted t h a t the maximum displacement occurs along the axis x = 0 and the displacement asymptot i - cally approaches zero as x --+ oo for a fixed value of time. Also the displacement profile does not approach a finite l imit for r --+ m. Physically this means t ha t once the plate is set in motion, the mot ion cont inues indefinitely. This is a result of the absence of a restoring force in the proposed viscous model. In this section, we shall discuss the previous results for a typicat projectile impac t : a ",
+            " projectile str iking a steel plate a t a velocity of 3000 fps: .! ;..~ oR 2 t = - - - r = 2.88 X 10-% sec. It v V0 cr~ = --~- <., = 1.37 X 106~ psi. (21) FoR2p w - ~ = 0.104~ in. r = Rx = O.15x in. Referring to Fig. 4, it can be seen t h a t the peaks in the shear stress occur in a very short t ime after impact : the maxinmm value of the stress for r = 1 (t = 2.88 X 10 -~ sec.) is approximately 123,000 psi, however, for r = 10 (t = 28.8 X 10 -6 sec.) the maximum stress is less than 7000 psi. Also, from Fig. 7 it can be seen t h a t for r = 1 (t = 2.88.10 - 6 Nov., ~963.J PLUG FORMATION IN PLATES 405 see.) the maxinmm displacement is approximately .05 in. Thus the viscous effects cause high stress (and small displacements) in the first few microseconds after impact. As t ime increases, the maximum stress decreases rapidly, and the displacement continues to increase (at a slower rate). The displacement profile for the velocity 2954 fps was measured from the photographs in Fig. 2 and plot ted on Fig. 7 convert ing to the nondimensional displacement ~ by means of the appropriate constants : p a n d u (Eq. 2 0 ) , R = .15 in. V0 = 2954fps. As shown in Fig. 7, the displacement was measured for three depths in the plate : (1) z = .08 in. (2) z = .31 in. and (3) z = .38 in. These profiles show tha t our assumption of radial symmet ry holds very well, however, there is a small variation of the displacement with plate depth. Qualitatively, the theoretical and experimental profiles have approximately the same shape in the region x < 1 ; outside this region, the experimental profiles approach the axis more rapidly than the theoretical. It should be remembered, however, that the predicted displacement profiles increase indefinitely with t ime and the experimental displacement profiles are the final displacements for the impact",
+            " The impact is represented as an initial velocity distr ibuted over a circular area of tile plate surface. The governing differential equation of motion Eq. 6, is the heat conduction equation. Hence, in the viscous mode employed here, the initial velocity disturbance does not propagate into the medium. On the contrary, the disturbance is felt at infinity immediately after the impact which means physically tha t an infinite amount of the plate is set in motion in the initial instant. Hence the curves for the shearing stress (Fig. 4), velocity (Fig. 6) and displacement (Fig. 7) are all asymptot ic to their respective zero ordinates, and show no fronts. The initial velocity disturbance is represented as a mathematical discontinuity (a step function), Eq. 1, and this discontinuity causes an infinite singularity in the initial shearing stress occurring at the boundary of the disturbance, as shown in Figs. 4 and 5. Figure 4 also shows that the maximum value of the shearing stress decreases rapidly with time. For the case of an initial velocity of 3000 fps incident to a steel plate, the 406 ANDREW PYTEL AND NORMAN DAVIDS [J",
+            " maximum stress decreases from a value of 123,000 psi to 7000 psi for the time interval 2.88-28.8 microseconds. It is also interesting to note from Fig. 4, that the infinite stress occurs at the boundary of the disturbance (that is, at x -- 1) and as the time increases, the position of the maximum stress moves further and further from the axis of the plate. By measuring the displacements in the photograph of a plugging experiment (Fig. 2) it was found that the assumptions of radial symmetry and independence of plate depth are approximately satisfied (Fig. 7). In the viscous model used here, the displacements gradually approach infinity as the time increases. To describe the problem realistically, the model would need to be extended to include a mechanism to stop the plate motion. Acknowledgment The authors gratefully acknowledge the support ,given this research by the Office of Ordnance Research, U. S. Army. Grateful acknowledgment is also given to the Watertown Arsenal Laboratory for permission to publish Fig. 2, which is photograph W.A. 639--4485"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002935_1.3629603-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002935_1.3629603-Figure1-1.png",
+        "caption": "Fig. 1 Geometric interpretat ion of cross-product technique",
+        "texts": [
+            "To) where the functions /<(x,, x2) and fi(xi, x2) satisfy the condition (9) i>ft(x 1, x2) Z>Mxu _ 0 dxi dx2 An integral of this last set of equations can still be found, and if the functions fi(xn xi) and fb(xi, x2) are properly selected, then some of the qualitative properties of the original set of equations (6) will be retained in equations (9). This integral can then be considered an appropriate candidate as a Liapunov function. Actually, the selection of appropriate functions fi(xi, x2) and Mxi) \u00a32) can be accomplished in the same operation that compares the original system (0) with the modified one (9). Fig. 1 illustrates this point. The vectors xc and xs represent the velocity vectors of equations (6) and (9), respectively. The cross product X So X Xi (10) represents a third vector which will be positive if the direction of tile vectors is as indicated in the figure, zero if the two vectors were to coincide, and negative if the vector xe were to point \"outward\" relative to the vector j-9. If the functions f-(xi, x2) and f-a(xi, x2) are selected so that the magnitude of this cross product is nonnegative, then conditions sufficient for the definitions of asymptotic stability to be fulfilled are satisfied"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002763_j.triboint.2011.06.007-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002763_j.triboint.2011.06.007-Figure1-1.png",
+        "caption": "Fig. 1. Thin fluid film squeezed between two parallel solid plates.",
+        "texts": [
+            " This to determine the model and the numeric method able to predict all the hydrodynamic characteristics of a thin lubricant film subjected to a periodic squeezing between conforming surfaces. This study also illustrates the lubricant film behaviour during an inversion phase of the applied dynamic load and rapid speeds change like those generated in the automobile engines crankshaft and connecting rods bearings or SFD. It contributes to the mastering of the performances of this type of dispositive. For the numerical study of lubricant film behaviour subject to periodic loading, we will consider two plates separated by a thin fluid film, as shown in Fig. 1. The top plate is oscillating while the bottom plate is stationary. It is considered also an orthogonal reference system, with the origin O, the axes Ox and Oz located on the bottom plate surface. The film thickness, measured in the direction of Oy axis, is a function of two spatial coordinates and time: h\u00bc h\u00f0x,z,t\u00de \u00f01\u00de Considering a thin oil film squeezed between two square and compliant two plates constantly parallel submerged in a lubricant bath. The top plate is oscillating harmonically along (y) while the bottom plate is stationary"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000724_s1064230707040077-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000724_s1064230707040077-Figure3-1.png",
+        "caption": "Fig. 3. The generalized linear element.",
+        "texts": [
+            " The simplified Jacobian matrix of the contact without friction is symmetric and negative semi-definite \u2206\u00b7 G f C\u03c1v T nT 0\u239d \u23a0 \u239b \u239e T ,= Gr N\u2206' C\u03c1v T Pn 0\u239d \u23a0 \u239c \u239f \u239b \u239e , G\u03c0 Gr\u03c1\u0303v\u2013 C\u03c1v T+= = \u00d7 N\u2013 \u2206' n\u03c9\u0303uv \u03c1\u0303v f \u03c1vnT n\u03c1v T\u2013( )\u239d \u23a0 \u239c \u239f \u239c \u239f \u239b \u239e , Gv N\u2206' C\u03c1v T Pn 0\u239d \u23a0 \u239c \u239f \u239b \u239e ,= G\u03c9 Gv \u03c1\u0303v .\u2013= \u00b7 \u00b7 N\u2206' N\u2206'\u00b7 JR* N\u2206' J*, JV* N\u2206' J*,= = \u00b7 J* Pn Pnr\u0303B\u2013 r\u0303BPn r\u0303BPnr\u0303B\u2013\u239d \u23a0 \u239c \u239f \u239b \u239e n r\u0303Bn\u239d \u23a0 \u239c \u239f \u239b \u239e nT nTr\u0303B\u2013\u239d \u23a0 \u239b \u239e .= = 572 JOURNAL OF COMPUTER AND SYSTEMS SCIENCES INTERNATIONAL Vol. 46 No. 4 2007 POGORELOV We use an element of this type, in particular, to simulate the strings when we do not take into account their inertial properties. We introduce two reference coordinate systems RSA and RSB that are stiffly connected to the bodies u and v and have the origins at the points A and B (Fig. 3). The axes of RSA and RSB are parallel to RSu and RSv. We assume that when the system moves, the displacement of RSB with respect to RSA is small. When this displacement takes place, the body v is subject to the force and the moment at the point B. Fu Mu \u239d \u23a0 \u239c \u239f \u239c \u239f \u239b \u239e G0 Cu\u2206Ru\u2013 F0 u M0 u \u239d \u23a0 \u239c \u239f \u239c \u239f \u239b \u239e = = \u2013 Crr u Cr\u03c0 u C\u03c0r u C\u03c0\u03c0 u \u239d \u23a0 \u239c \u239f \u239c \u239f \u239b \u239e \u2206ru \u2206\u03c0u \u239d \u23a0 \u239c \u239f \u239c \u239f \u239b \u239e . Here, Cu is a constant matrix of dimension 6 \u00d7 6, G0 is the value of the force and moment when RSA and RSB coincide, and \u2206r and \u2206\u03c0 are the displacement vectors of the point B and the small rotation of RSB with respect to RSA"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003174_robot.2010.5509512-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003174_robot.2010.5509512-Figure2-1.png",
+        "caption": "Fig. 2. Gantry-Tau schema with notation for variables and parameters",
+        "texts": [
+            " Their placement on plate and carts according to the so-called Tau configuration is such that the links belonging to one cluster form parallelograms, which assures a constant end- 978-1-4244-5040-4/10/$26.00 \u00a92010 IEEE 3709 effector orientation. The Gantry-Tau robot has thus three purely translational DOFs. A detailed solution of the kinematics problem can be found in [2] or [6]. Provided that the end-effector orientation is constant due to a perfect spherical joint placement, it is sufficient to consider the simplified robot shown in Fig. 3. The closure equation for link i is then (notation see Fig. 2): L2 i \u2212\u2016 T \u2212 sCi \u2016 2 2 = 0 (1) The cart position sCi of the simplified model can be expressed as: sCi = sC0 i + qi \u00b7 vi (2) where vi is the unit vector in positive track i direction. The track offset sC0 i of the simplified model is sC0 i = C0 i \u2212 RT \u00b7Pi (3) where Pi is link i\u2019s spherical joint position on the end-effector plate expressed in TCP coordinates and RT the rotation matrix between the TCP and the global frame. The nominal kinematic model assumes perfectly linear actuators and constant end-effector orientation guaranteed by the Tau-configuration of the spherical joints",
+            " The cart position Ci for the commanded actuator position qi is now interpolated linearly between the two cart measurements mC k i and mC k+1 i whose corresponding actuator positions qk i and qk+1 i are closest to qi: Ci (qi) = mC k i + qi \u2212qk i qk+1 i \u2212qk i ( mC k+1 i \u2212 mC k i ) (4) Ci in Eq. 4 is, unlike Ci in Eq. 2, expressed in the coordinate system used for measurement set 2. Instead of optimizing the track direction vi and offset sC0 i as for the nominal model, the coordinate frame transformation between track and TCP measurement frame has to be calibrated. For the kinematic error model of the arm structure, all 6 links have been taken into account (see Fig. 2) as well as the TCP orientation errors that arise if the links in one cluster have slightly different lengths or if the joint placement on carts and end-effector is not according to the Tau configuration. The closure equation for link i is then: L2 i \u2212\u2016 T + RT \u00b7Pi \u2212 Ci \u2016 2 2 = 0 . (5) To evaluate the kinematic modeling in the previous section, the calibration results of different combinations of actuator and arm structure models are compared. Fig. 7 illustrates the different models: Model 1 is the nominal kinematics with linear actuators and a reduced arm structure",
+            " Within the given range of [-650,700] mm for the actuators, the modified track directions vi result in the smallest TCP positioning changes, while the rotation matrix of the transformation between the 2 different laser tracker positions gives larger variations, as the distance from the initial cart position magnifies the 5\u00b710\u22125 change. For both the joint offsets on carts and end-effector plate the cost function is much more sensitive to the x component than to the y and z components (coordinate system see in Fig. 2). The link lengths Li (see Fig. 9 for L1) result in the largest sensitivity. Considering Fig. 5, it appears reasonable that cart 1 gains the least by a piecewise linear instead of a linear actuator 0 10 20 30 40 50 60 \u22120.04 \u22120.02 0 0.02 0.04 angerr 0 10 20 30 40 50 60 \u22120.04 \u22120.02 0 0.02 0.04 0 10 20 30 40 50 60 \u22120.04 \u22120.02 0 0.02 Validation point index \u03b1 [d eg ] \u03b2 [d eg ] \u03b3 [d eg ] Fig. 10. Changed orientation errors for increasing (red) and decreasing (blue) Px,3 about 50 \u00b5m. modeling, while improved results can be expected for cart 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003997_s00521-012-1163-3-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003997_s00521-012-1163-3-Figure2-1.png",
+        "caption": "Fig. 2 Path vector set regeneration a path vector set P1, b path vector set P2",
+        "texts": [
+            " When the agent reaches position q1, there are obstacle points meeting Rule 1. The intermediate point qinsert1 is then calculated in accordance with the Rule 2, and the path is regenerated. Agent moves along the regenerated path, repeating to detect the environment and calculate the obstacle point with Rule 1. When it reaches the position qinsert1 and turns to the next path direction, there are obstacle points meet Rule 1. The intermediate point qinsert2 is calculated, and the path is regenerated again. 2.4 Path regeneration As shown in Fig. 2a, the initial path vector set P0 only contains one vector formed by the line connecting start position qstart with goal position qgoal. The angle of the vector is zero. (If the agent is not facing the goal position, steer the agent to the goal position in its original place.) P0 \u00bc a01 \u00f02\u00de a01 \u00bc 0; a01k k\u00f0 \u00de \u00f03\u00de Assuming Obstacle1 blocked the path, we add intermediate point qinsert1 into initial path vector set P0 to get the regenerated path vector set P1, P1 \u00bc a11; a12 \u00f04\u00de a11 \u00bc h11; a11k k\u00f0 \u00de \u00f05\u00de a12 \u00bc h12; a12k k\u00f0 \u00de \u00f06\u00de where a11 and a01 are known. a12 and h12 can be calculated based on the Cosine theorem, a12k k \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi a01k k2\u00fe a11k k2 2 a01k k a11k k cos h11 q \u00f07\u00de h12 \u00bc cos 1 a01k k cos h11 a11k k\u00f0 \u00de= a12k k\u00f0 \u00de \u00f08\u00de As shown in Fig. 2b, the agent keeps moving in accordance with path vector set P1. During the moving process along path vector a12, Obstacle2 will block the path. We add intermediate point qinsert2 into the path vector set P1 to get the regenerated path vector set P2, P2 \u00bc a21; a22; a23; a24 \u00f09\u00de a2k \u00bc \u00f0h2k; a2kk k\u00de; k \u00bc 1; 2; 3; 4 \u00f010\u00de where h22 = h12, a23, and a22k kis known. a24 and h24 can be calculated based on the Cosines theorem, a24k k \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi a12k k a22k k\u00f0 \u00de2\u00fe a23k k2 2 a12k k a22k k\u00f0 \u00de a23k k cos h23 q \u00f011\u00de h24 \u00bc cos 1 a12k k a22k k\u00f0 \u00de cos h23 a23k k\u00f0 \u00de= a24k k\u00f0 \u00de \u00f012\u00de Similarly, we add intermediate point qinserti into path vector set Pi to get regenerated path vector set Pi11, Pi\u00fe1 \u00bc a\u00f0i\u00fe1\u00de1; a\u00f0i\u00fe1\u00de2; "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001199_0278364907085565-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001199_0278364907085565-Figure1-1.png",
+        "caption": "Fig. 1. Generic non-modular manipulator.",
+        "texts": [
+            " As the computationally distributed control architecture proposed here has been devised by \u201cdistributing\u201d the computations required for controlling a conventional (i.e. not modular) manipulator among different processing units, it is convenient to start by first briefly recalling the basic features of the kinematic-based centralized controller typically employed for governing the operational space motion of robotic manipulators. To this aim consider a generic manipulator, for instance similar (without loss of generality) to the commonly used manipulator with seven degrees of freedom sketched in Figure 1, where the wrist is constituted by a rotational joint (typically of Euler or roll\u2013pitch\u2013yaw type) with three degrees of freedom. In Figure 1, frame g represents the \u201cgoal frame\u201d, which has to be reached (in position and orientation) by the so-called \u201ctool frame\u201d t , rigidly attached to the \u201cend-effector frame\u201d e . Positions (described through vectors p 3) and attitudes (described through rotational matrices R 3x3) of frames g and t with respect to an inertial \u201cworld frame\u201d w are respectively embedded within the transformation matrices g T t g R t pg t 01x3 1 t T q t R q pt q 01x3 1 (1) The former is given as an external time-varying reference input, while the latter, depending on actual posture, is real-time computed as the result of the product t T 0 T 0 e T q et T (2) where e t T and o T (the transformation matrices of t with respect to e and of 0 with respect to w , respectively) are given as external configuration data, while o e T q (the transformation matrix of e with respect to 0 ) is real-time evaluated by means of the knowledge of robot geometry and of the current joint position vector q",
+            " More specifically, for the case at hand, the task space is 2 and the subspaces on which VPU 1 and VPU 2 project their input errors are those orthogonal to h1 and h2, respectively. As a consequence, although the conditions guaranteeing the convergence of the proposed technique are analyzed more formally in Section 3.3, it can already be reasonably argued that the proposed algorithm always assures the non-increasing behavior of the norm of the error vectors. With the above preliminary considerations in mind, we shift attention back to the original and more general case of a 3D manipulator (such as that depicted in Figure 1) whose tool frame is asked to accomplish a generic six-component velocity task x . This general case differs from the special case dealt with in the previous section in two main ways: the enlarged task dimension (six instead of two) and the increased number of at CARLETON UNIV on March 16, 2015ijr.sagepub.comDownloaded from VPUs (equal to the number of degrees of freedom, now assumed to be greater than six). As will be pointed out in the following, despite such differences, the problem of finding a distributable kinematic inversion procedure can be stated and solved by adopting an approach that is very similar to that used previously"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure4-1.png",
+        "caption": "Fig. 4. Substructure A+B: constrained BIW for bending stiffness",
+        "texts": [
+            " (1) Equation (1) indicates how the system matrix of the coupled system HC can be calculated from the system matrices of the components (HA and HB) The indices are related to input, output and coupling points as indicated in Fig. 1. The FBS technique can be used to convert the free-free system to a constrained system: \u2022 subsystem A: free-free FRF's of BIW as shown in Fig. 2 \u2022 subsystem B: ground To represent the static test bench condition for torsional stiffness, the rear domes will be grounded (Fig. 3). To represent the static test bench condition for bending stiffness, rear and front domes will be grounded (Fig. 4). Forced Response The bending stiffness of a vehicle body is measured by clamping the body at the four domes and applying a load at the 4 seat bolting positions (F1 to F4) as is represented in Fig. 5. (2) After calculating the coupled system matrix Hc from equation (1) the displacement in the output points d(\u03c9) can be easily calculated using a forced response described in equation (2). This is a general approach which could be used for several load cases. The forced response can be done over a certain frequency range (e"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001304_j.snb.2009.08.048-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001304_j.snb.2009.08.048-Figure4-1.png",
+        "caption": "Fig. 4. Schematic arrangement of the optical components and optical path.",
+        "texts": [
+            " The objective of the instrument development was to measure his photon flux with a higher accuracy than the estimated S/B atio. For optimal acquisition of the photon flux this measurement equires a body part without fur that could falsify the measurement. or that reason, the tail of the mouse, which is well vascularized, as used. In order to focus the radiation created by the mouse tail, the optial path has to be adapted to the geometry of the tail. The geometry f the mouse tail is assumed as cylindrical. For optimal detection, the mouse tail is located along the focus ine of a cylindrical ellipsoid mirror (Fig. 4, right hand side). In the lane perpendicular to the mouse tail, the radiation is focused by he ellipsoid mirror which is also the focus of a cylindrical planeonvex ZnSe lens. After passing this lens, the radiation is converted nto a well-defined beam that includes most of the power within the iameter of the detector. The radiation leaving the tail tangential or arallel is reflected by the mirror and then focused by a system of wo cylindrical plan-convex lenses into the detector (Fig. 4 left hand ide, showed as a square). This arrangement of optics is intended to ptimize the detected power of the infrared radiation in the detecor, it is not a reproduction of the original geometry of the mouse ators B 142 (2009) 502\u2013508 tail in the detector. The signal represents rather an average of the radiant area. Simulations with this optical arrangement have a collection efficiency of 4.87% for a detection area of 5 mm \u00d7 5 mm translating the calculated photon flux above into a flux of 4 \u00d7 1013 s\u22121 at each detector (2 mm \u00d7 2 mm, assuming a radiating area of the mouse tail of 471 mm2)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001107_iet-cta:20070072-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001107_iet-cta:20070072-Figure1-1.png",
+        "caption": "Figure 1 Illustration on region control of UVMS",
+        "texts": [
+            " It is also important to note that when the desired region is very small, it reduces to a point. Control Theory Appl., 2008, Vol. 2, No. 9, pp. 819\u2013828 10.1049/iet-cta:20070072 Recently, a region-reaching control scheme [11] is proposed for robot manipulator. In this new control concept, the desired objective can be specified as a region instead of a point. However, the results in [11] are limited to robot manipulator where a single desired region is specified for the end effector as illustrated in Fig. 1a. Besides reaching tasks, region-reaching concept is also useful for control of macro/mini-structures, where a large region (secondary region) is specified for the macro-system and a smaller region (primary region) is specified for the mini-system. That is, a large region should be specified for the macro-system instead of restricting its position to a point as shown in Fig. 1b. This gives the macro-system more freedom to adjust its position, whereas the minisystem is performing various tasks. In this paper, we proposed a new region-reaching controller for an underwater vehicle (macro-system) mounted with a manipulator (mini-system). Many research efforts have been devoted to the development of underwater robotics [12\u201320], as the need for exploring and preserving the oceanic environments has gained significant momentum. In the conventional setpoint control problems of underwater vehicles [12], the desired position is specified as a point",
+            " However, in some applications of underwater vehicles, the control objective is specified as a region instead of a point, for example, maintaining the underwater vehicle within a minimum and maximum depth in water; underwater vehicle travelling inside the pipeline for specific 819 & The Institution of Engineering and Technology 2008 820 & T task; and avoiding an obstacle located at a specified region. This paper presents a region-reaching control concept for underwater vehicle-manipulator systems (UVMS), where the desired objectives can be specified by regions instead of points. For UVMS, two desired regions are being specified, namely primary region and secondary regions, as illustrated in Fig. 1b. The proposed region control concept is also a generalisation of setpoint control problem because when the desired region is specified arbitrarily small, the control objective reduces to a point. 2 Dynamics In this section, the structure and properties of UVMS kinematics and dynamics are briefly reviewed. An underwater vehicle with an n-link manipulator attached on it is considered. Two common vectors that being used in defining the underwater vehicle state vector are h and v. The vector h is defined as h \u00bc [hT 1 hT 2 ]T, where h1 \u00bc [x y z]T is the vehicle position vector in the earth fixed frame and h2 \u00bc [f u c] T is the vehicle Euler angle in the earth fixed frame",
+            "1049/iet-cta:20070072 IET doi or moment acting on the vehicle as well as joint torques. Several important properties of the UVMS dynamics equation described in (7) are [16]: Property 1: The inertia matrix, M(q) including the added mass is symmetric and positive definite. Property 2: The matrix, _M(q) 2C(q, z) is skewsymmetric. Property 3: The hydrodynamic damping matrix, D(q, z) is strictly positive such that D(q, z) . 0. 3 Region control law for UVMS For UVMS, two desired regions are being specified, namely primary and secondary regions. As illustrated in Fig. 1b, a primary region is specified for the manipulator end effect and a secondary region is specified for an underwater vehicle. Usually, a bigger region is specified for the secondary region to give the underwater vehicle more freedom while the manipulator is performing various tasks. The desired region for underwater vehicle, which is the secondary region, can be specified as fV (dho) \u00bc fV1 (dho1 ) fV2 (dho2 ) .. . fVN2 (dhoN2 ) 2 666664 3 777775 0 (8) where dhoj \u00bc h hoj , hoj is the reference point of the jth desired region, j \u00bc 1, 2, "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure9.7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure9.7-1.png",
+        "caption": "Fig. 9.7-2 Lateral vibration of beam with axial tension [9.2, 9.7].",
+        "texts": [
+            " (1) The Effect of Rotary Inertia and Shearing Deformation When the cross-section dimensions of a beam are suffi ciently large so that shear forces (forces tangent to the cross section) produce rotation in the cross section comparable to rotation produced by bending moments, the beam is sometimes called a Timoshenko beam [9.2]. With such beams, the cross section may signifi cantly resist rotation during vibration due to the moment of inertia of a cross-sectional beam element [9.2, 9.7]. To explore this, consider a beam vibrating laterally in the X-Y plane\u2014that is, where the X-axis is along the beam axis and the displacement is in the Y direction. Consider an element of the beam, also in the X-Y plane, as represented in Fig. 9.7-1. Th e positive directions for shear forces, bending moments, and displacement are those shown in the fi gure. VibrationAnalysis_txt.indb 336 11/24/10 11:50:06 AM Continuous Systems | Chapter 9 337 Th e slope of the defl ection curve of the beam is \u2202 \u2202u x/ , which depends not only on the rotation of cross sections of the beam but also on the shear. Let be the slope of the defl ection curve when the shear force is neglected and let be the angle of shear at the neutral axis in the same section. Th en we have \u2202 \u2202 = \u2212u x (9"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002135_s0005117910110068-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002135_s0005117910110068-Figure1-1.png",
+        "caption": "Fig. 1. Flipped pendulum with wheel.",
+        "texts": [
+            " 11 2010 Reduction in \u03be leads to a more rigid constraint on the diagonal elements and the eigenvalues of the matrix A22 + CA12 and guarantees that the condition diag(A22 + CA12) \u03be is satisfied. By selecting different parameters \u03bc and \u03be and solving the system of linear matrix inequalities (29), we obtain from (26) different switching surfaces for which one should verify additionally condition (19). At seeking the switching surface for the R\u03b5-stabilization, only the constraint on the parameter \u03be is varied: \u03be ln 2 2h0 . 5. NUMERICAL EXAMPLE Let us consider the problem of stabilization of the wheel-controlled flipped pendulum (see Fig. 1). The wheel is fixed on the axle of a DC electric motor located at the free end of the pendulum. The corresponding linearized system of motion equations [18] is as follows { J\u03c8\u0308 + (Jr + Jf )\u03c9\u0307 = (Md+ml)g\u03c8 (Jr + Jf )(\u03c8\u0308 + \u03c9\u0307) = c1u\u2212 c2\u03c9, (30) where \u03c8 is the angle of pendulum deviation, \u03c9 is the angular speed of the wheel, J = Jp +ml2 + Jr + Jf , Jp = 1.2\u00d7 10\u22123 [kg m2] is the pendulum moment of inertia, Jr = 1.2\u00d7 10\u22126 [kg m2] is the rotor moment of inertia, Jf = 7.65 \u00d7 10\u22125 [kg m2] is the wheel moment of inertia, M = 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003828_s11431-013-5313-3-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003828_s11431-013-5313-3-Figure2-1.png",
+        "caption": "Figure 2 Computational mesh for Eckardt impeller.",
+        "texts": [
+            " The results indicate that the fine mesh which has 800 k cells already provides a mesh-independent solution and this mesh was then used for the rest of this study. The fine mesh distribution is 200\u00d770\u00d7 50 streamwise, spanwise and blade-blade, respectively. In this mesh 22 points are used in the tip gap region from blade tip to the casing and 20 points across the blade tip in the pitchwise direction. The value of Y+ was found to be typically between 10 and 40 on all walls, which is appropriate for the standard wall function used in the CFD model. The computational mesh used for the study is shown in Figure 2, together with the scaled-up of the mesh in the tip clearance area. The computational domain consists of an unshroud impeller followed by a vaneless diffuser. A frozen-rotor interface is used between the rotating and stationary components. The mesh has a structured H-O topology and consists of approximately 800 k nodes. The computations were performed with the nominal clearances as shown in Table 1. Additional computations were performed with zero tip clearance with the shroud stationary so that the scraping effect could be modeled"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001752_s10957-010-9722-1-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001752_s10957-010-9722-1-Figure3-1.png",
+        "caption": "Fig. 3 Piecewise linear function \u03c8(\u03c4)",
+        "texts": [
+            " (17) Here, \u03b8 is a parameter lying within the interval ]0,1/2[, while b1 and b2 are positive constants. Function \u03c8(\u03c4) from (17) satisfies the condition imposed in (10). This function must be continuous for all \u03c4 . To satisfy this property, it is necessary and sufficient to impose the continuity requirement at the instant \u03c4 = \u03b8 . According to (17), this condition implies the following equality b1\u03b8 = b2(1/2 \u2212 \u03b8) (18) imposed on parameters b1, b2, and \u03b8 . The graph of function \u03c8(\u03c4) defined by (17) is shown in Fig. 3. Substituting function \u03c8(\u03c4) from (17) into integrals I1 and I0 from (15) and calculating these integrals, we obtain I1 = 2b1b2 b1 + b2 , I0 = b2 1b 2 2(b1 \u2212 b2) 4(b1 + b2)2 . (19) The condition I0 < 0 is satisfied, if and only if b1 < b2. In this case, the quadratic trinomial F(w) from (14) has a unique positive root w\u2217 that defines the asymptotically stable stationary velocity of the system. Inserting expressions given by (19) into (16), we obtain w\u2217 = b1b2 b1 + b2 (\u221a 1 + b2 \u2212 b1 4 \u2212 1 ) > 0. (20) This solution is expressed through non-dimensional variables introduced in (6)\u2013(8)",
+            " Suppose that the function \u03c8(\u03c4) increases monotonically from 0 to \u03c80 on the interval ]0, \u03b8 [, reaches its maximum \u03c8(\u03b8) = \u03c80 at \u03c4 = \u03b8 , and decreases monotonically from \u03c80 to 0 on the interval ]\u03b8,1/2[. Here, \u03b8 \u2208]0,1/2[ and \u03c80 > 0. The behavior of \u03c8(\u03c4) on the interval ]1/2,1[ is determined by (10). Thus, we assume that the optimal motion of the tail consists of the deflection phase, where |\u03c8 | grows monotonically, and of the retrieval phase, where |\u03c8 | decreases monotonically. This behavior of \u03c8(\u03c4) and \u03d5(t) is qualitatively similar to the one shown in Fig. 3. Under the assumptions made, we can evaluate the integrals defined by (15) as follows: I1 = 2 \u222b \u03b8 0 d\u03c8 d\u03c41 d\u03c41 \u2212 2 \u222b 1/2 \u03b8 d\u03c8 d\u03c41 = 4\u03c80, I0 = 2(I \u2032 \u2212 I \u2032\u2032), I \u2032 = \u222b \u03b8 0 \u03c8 ( d\u03c8 d\u03c41 )2 d\u03c41, (29) I \u2032\u2032 = \u222b 1/2 \u03b8 \u03c8 ( d\u03c8 d\u03c41 )2 d\u03c41. We will choose the time-history of the angle \u03d5(t), so that to maximize the average stationary velocity v\u2217 defined by (8) and (12): v\u2217 = (\u03bca/T0)w \u2217, (30) where w\u2217 is given by (16). To do that, we will first fix the constant parameters T (or T0 = T/\u03bc), \u03b8 \u2208]0,1/2[, and \u03d50 > 0 (or \u03c80 = \u03d50/\u03bc), and maximize the non-dimensional velocity w\u2217 by choosing the function \u03c8(\u03c4) or its derivative d\u03c8/d\u03c4 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001883_icndc.2010.72-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001883_icndc.2010.72-Figure3-1.png",
+        "caption": "Figure 3. Robot\u2019s foot and ground surface, Spring-damping model of robot\u2019s foot and ground surface",
+        "texts": [
+            " His leg muscles will be relax to absorb the shock when landing and be hardened to maintain the balance after landing. Such elastic features of leg in walking takes our attention. We can imitate the human\u2019s leg to realize biped robot\u2019s variety locomotion. Because the detailed model of the environment is unavailable in most practical situation. Thus, we select the compliance control to deal with the impact force. Suppose that the robot foot is connected to the ground with a spring and damper shown in Fig. 3. Force torque(F/T) sensor has been mounted between the foot and the ankle to measure the impact force. Also, an ankle coordinate frame is fixed on the center of the support leg\u2019s ankle. The landing force between the landing foot and the ground, F \u2208 R6, can be given by: Mp\u0308e + Dp\u0307e + Kpe = F pe = p \u2212 pd (1) where M \u2208 R6\u00d76 is the robot mass matrix. D \u2208 R6\u00d76 and K \u2208 R6\u00d76 are the damping and stiffness matrices, respectively. p \u2208 R6 and pd \u2208 R6 denote the actual and desired position vectors of foot. The angle of each joint can be measured by encoder attached to the joint"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003727_j.engfailanal.2011.06.004-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003727_j.engfailanal.2011.06.004-Figure8-1.png",
+        "caption": "Fig. 8. Position of the critical nodes that lie on the lateral surface of the second U-bolt.",
+        "texts": [
+            " (11), the fully corrected endurance limit of the U-bolt is estimated as Se = 120.9 MPa. After computation of principal alternating and mean stresses for all nodes, the effective alternating, Sae, and effective mean stress, Sme, are obtained by utilizing Eqs. (7) and (8). By using the modified Goodman approach (Eq. (5)) and applying fatigue stress concentration factor, Kf, fatigue stress, SNf, is obtained for all nodes. The fatigue stress for six critical nodes is summarized in Table 2. The position of these nodes is illustrated in Fig. 8. Among all critical nodes, node 6416 has the maximum value of SNf = 155.2 MPa. Fatigue life for this critical node is obtained as 3.63 million cycles or 593 h using Basquin\u2019s equation (Eq. (6)). It is noted that the position of fracture is dictated by the maximum fatigue stress (SNf) that is near the highest stress regions (Figs. 7 and 8). Graphical representation of the fatigue stress distribution is impossible with the software used in this study. Therefore, distribution of the maximum and minimum of the largest principal stress at fracture region is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002856_iecon.2011.6119562-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002856_iecon.2011.6119562-Figure3-1.png",
+        "caption": "Fig. 3 Influence of yoffset and ymod on the transient current change for a given pulse sequence in phase V. (asymmetry position according to Fig. 2)",
+        "texts": [
+            " 2 the above described coherences are depicted for an asymmetry caused by a missing slot wedge in the stator at angular position \u03b3=50\u00b0. The resulting asymmetry is assumed to lead to an ideal sinusoidal modulation of the leakage inductance. The corresponding direction dependence of the leakage inductance is indicated as ellipse with the minimum inductance in the direction of the magnetic axis corresponding to a magnetomotive force located in the slot with the missing wedge. The same ellipse is obviously obtained with a wedge missing at \u03b3=230\u00b0 as already considered in (3),(6). The corresponding signals are depicted in Fig. 3 for a difference voltage phasor vS,I-II pointing in direction of phase V (vS,I=+V and vS,II=-V). The resulting current change phasor \u0394is,I-II/\u0394\u03c4 is composed of two parts already described, depicted as black-dashed arrows denoted vS,I-II\u00b7ymod and vS,I-II\u00b7yoffset according to (4) and (5). The position of the asymmetry is chosen according to Fig. 2 with the main axis of the ellipse denoted \u201cmaximum inductance\u201d. The angle \u03b4 defines the difference of the direction of the maximum transient inductance (slot wedge) with respect to the direction of the difference voltage vector (vS,I-II) of the pulse excitation",
+            " In the control and measurement setup a processor takes care of the voltage pulses for excitation and the trigger signals for measurement. The voltage pulse sequence is generated from stored values. Thus the voltage phasor (vS,I-II) can be calculated in advance. The current difference phasor (\u0394is,I-II/\u0394\u03c4) is resulting from current samples taken at specific instances, like shown in Fig. 1. As this current difference phasor is composed of a symmetrical part as well as an asymmetrical portion, it is necessary to clearly separate these two shares in the measured signal. This can be done as follows. As depicted in Fig. 3 the yoffset influenced part of the current difference phasor points in the direction of the excitation pulses. The voltage pulse sequences are applied to the three main phase directions subsequently. Combining the three resulting current difference phasors by adding them together leads to only one current difference phasor. In this phasor calculation the symmetrical shares are eliminated as zero sequence value. This phasor now contains only information on the machine\u2019s asymmetries. In the following this overall current difference phasor is thus denoted as asymmetry phasor"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001635_9783527632534.ch8-Figure8.6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001635_9783527632534.ch8-Figure8.6-1.png",
+        "caption": "Figure 8.6 Schematic representation of the biocatalytic polymerization of aniline in presence of glucose oxidase, ( Reproduced with permission from [57] . Copyright \u00a9 (2007) Elsevier ).",
+        "texts": [
+            " The use of crude extracts, instead of the highly purifi ed enzyme, is a more economic alternative and therefore attractive from the industrial point of view. 8.5.3 Synthesis of PANI Using Other Enzymes Glucose oxidase is an oxidoreductase that catalyzes the oxidation of \u03b2 - d - glucose into d - glucono - 1,5 - lactone, which then is hydrolyzed in aqueous media to gluconic acid, while reducing molecular oxygen to hydrogen peroxide. It is a very important enzyme, which is usually employed in glucose biosensors. Ramanavicius et al . [57] took advantage of the enzymatically produced hydrogen peroxide to oxidize aniline in a broad pH conditions (Figure 8.6 ), ranging from 2.0 to 9.0. Optimum conditions for hydrogen peroxide production were close to physiological conditions, although is highly likely that PANI synthesized under this conditions consists primarily of oligomers and ortho - linked/crosslinked structures. Longoria et al . [58] recently reported the oxidation of highly chlorinated aniline by Chloroperoxidase from Calderomyces fumago . The specifi c activity shown by this enzyme was two orders of magnitude higher than that of HRP. Polymeric insoluble products were reported as the main material obtained from the reaction, reaching typically between 87 and 95% of the transformed compounds"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001178_j.scriptamat.2007.08.039-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001178_j.scriptamat.2007.08.039-Figure1-1.png",
+        "caption": "Figure 1. Single-walled nanotube (liquid drop) geometry, under large contact/deformation (a), squashed configuration (b), and continuum (up) or atomistic (down) computed shapes for a (20,0) carbon nanotube [3].",
+        "texts": [
+            " Such reasons have motivated the present study, with the aim of providing simple laws for treating this complex phenomenon. The adhesion of a small (i.e. one for which surface tension prevails over gravity) liquid drop is fully described by the contact angle h (a function of the 1359-6462/$ - see front matter 2007 Acta Materialia Inc. Published by El doi:10.1016/j.scriptamat.2007.08.039 * Tel.: +39 115644895; fax: +39 115644899; e-mail: nicola.pugno@ polito.it liquid/solid/vapour surface energies) between drop and substrate: see Figure 1a (with r = 0). With R0 the radius of the drop in air and R the radius of curvature of the spherical cap describing the adhering drop, the radius of the contact area a can be calculated assuming mass conservation. The adhesion between single-walled nanotubes of radius R0 can be similarly described by the contact angle h (which we expect to be a function of the adhesion work and bending stiffness), the radius of curvature R of the deformed non-contact segment, the radius of curvature r of the blunt notches (that, as a first approximation, could be assumed to be zero, as for a liquid drop) and the contact half-length a = (R r)sinh (see Fig. 1a). The nanotube mass conservation basically imposes its inextensible condition, i.e. a + (p h)r + Rh = pR0 (Fig. 1a). Accordingly we deduce: a pR0 \u00bc 1 r=R0 1\u00fe h= sin h : \u00f01\u00de For small contacts h! p and a/(p R0) (1 r/R0)sinh/ h. This asymptotic solution in the limit of r/R0! 0 can be directly compared with the analytical result reported in Ref. [2]; thus, we can define the contact angle for a single-walled nanotube, having Young\u2019s modulus E, Poisson\u2019s ratio v, thickness t and contact surface energy cS (here due to van der Waals attraction): sin h h \u00bc 1 R 0 R0 ; R 0 \u00bc ffiffiffiffiffiffiffi D 2cS s ; D \u00bc Et3 12\u00f01 v2\u00de ; \u00f02\u00de sevier Ltd",
+            "40 J/m2) radius r of the blunt notches (and thus, in our treatment, 0 6 r=R0;R 0=R0 6 1\u00de. Introducing Eq. (2) into Eq. (1) we find the following nonlinear law: a pR0 \u00bc \u00f01 R 0=R0\u00de\u00f01 r=R0\u00de 2 R 0=R0 : \u00f03\u00de For small contacts/deformations (h! p or R 0=R0 ! 1) and vanishing blunt radius (r/R0 ! 0) the prediction of Eq. (3) is identical to the asymptotic solution reported in Ref. [2], whereas for large contacts/deformations (h! 0 or R 0=R0 ! 0) a=\u00f0pR0\u00de \u00bc \u00f01 r=R0\u00de=2, as coherently imposed by the inextensible condition (see Fig. 1b). However, note that in the limit of small contacts/deformations and non-vanishing blunt radius our prediction is slightly different from that reported in Ref. [2]; for example for r \u00bc R 0, a=\u00f0pR0\u00de \u00bc \u00f01 R 0=R0\u00dea with a = 2, whereas in Ref. [2] a = 1. The (maximum) height of the flattened nanotube can be geometrically derived as (Fig. 1a): h 2R0 \u00bc R\u00fe r 2R0 R r 2R0 cos h; R pR0 \u00bc 1 a pR0 p h p r R0 h ; \u00f04\u00de where h is defined in Eq. (2). For h! 0, h,/(2R0)! r/R0 as coherently imposed by the inextensible condition (see Fig. 1b), whereas for h! p, h,/(2R0) ! 1. During the loss of adhesion, the classical fracture mechanics energy balance must hold. Accordingly, the opposite of the variation of the total potential energy (elastic energy minus external work) with respect to the crack surface area (complementary to the contact area) must be equal to the work of adhesion 2cS (see Refs. [7,8]). Thus dU = 4cSda (the external work is here zero), where U denotes the elastic energy per unit length stored in the nanotube. By integration, following Ref. [9], we can calculate the energy stored in the largely deformed nanotube: U \u00bc 4cSa \u00bc 4pR0cS \u00f01 R 0=R0\u00de\u00f01 r=R0\u00de 2 R 0=R0 : \u00f05\u00de Let us consider the numerical example investigated in Ref. [2], as shown in Figure 1c (adapted from Ref. [2]). Note the similarity between the deformed shapes of a nanotube and a drop. For the investigated (20,0) nanotube (R0 = 7.83 A\u030a) the size of the contact calculated [2] by molecular simulations is 2a = 8.73 A\u030a, with a flattened nanotube height equal to h = 11.75 A\u030a (Fig. 1c). Such values are comparable with those obtained by numerically integrating the elastica differential equation, which yields [2] 2a = 7.67 A\u030a and h = 12.96 A\u030a (Fig. 1c). The critical radius R 0 must be around 5 A\u030a, as reported in Ref. [2] \u00f0R 0 4:77 A\u030a\u00de or as suggested by the shape of self-collapsed nanotubes [3] \u00f0R 0 r 5:5 A\u030a\u00de. Note that, taking the theoretical value of 2cS = 0.40 J/m2, R 0 4:77 A\u030a would correspond to D = 0.91 \u00b7 10 19 Nm, whereas R 0 5:5 A\u030a to D = 1.21 \u00b7 10 19 Nm (in Ref. [3] values from the literature between D = 1.37 \u00b7 10 19 Nm and D = 2.35 \u00b7 10 19 Nm are reported, whereas in Ref. [2] slightly smaller values, as emphasized by the same authors, were calculated)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003382_978-1-84996-062-5_5-Figure4.12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003382_978-1-84996-062-5_5-Figure4.12-1.png",
+        "caption": "Figure 4.12 Themain process parameters",
+        "texts": [
+            " They found that the CO2 laser plume had the finer particles, with 93% below 100 nm and only 2% above 1 \u03bcm, whereas the Nd:YAG laser vapour had only 78% below 100 nm but 14% above 1 \u03bcm. Scattering by bothRayleigh andMie scatteringwas calculated to be significant for Nd:YAG laser processing, whereas inverse bremsstrahlung absorption was higher with the CO2 laser. Certainly Nd:YAG laser welding generates considerable dust compared with CO2 laser welding. Themain process parameters are illustrated in Figure 4.12. They are as follows: \u2022 Beam properties; power, pulsed or continuous; spot size and mode; polarisation; wavelength. \u2022 Transport properties: speed; focal position; joint geometries; gap tolerance. \u2022 Shroud gas properties: composition; shroud design; pressure/velocity. \u2022 Material properties: composition; surface condition. There are two main problems in welding: lack of penetration and the inverse \u201cdropout\u201d. These are the boundaries for a good weldmade with a given power as illustrated in Figure 4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001141_j.jmatprotec.2008.02.071-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001141_j.jmatprotec.2008.02.071-Figure9-1.png",
+        "caption": "Fig. 9 \u2013 Relationship between clearance angles E and at point E.",
+        "texts": [
+            " 8, the normal vector of the involute curve at point B is tangent to the base circle. The directional vector \u2212\u2192 OG is thus the same as the tangential vector T\u2032 B at point B. The pressure angle at point B, i.e., B is thus obtained as follows: rB cos B = rb (52) where rB is the position vector of point B and rb is the radius of the base circle. As vectors TB and T\u2032 B are perpendicular to \u2212\u2192 BO and \u2212\u2192 BG, respectively, the relief angle B is obtained as B = 2 \u2212 B (53) 4.2.3. Analysis of the clearance angle E In Fig. 9, TE is the tangential vector at point E while T\u2032 E is the vector perpendicular to the position vector rE. Applying the theory of involutometry, the clearance angle E is obtained as 854 j o u r n a l o f m a t e r i a l s p r o c e s s i n g t e c h n o l o g y 2 0 9 ( 2 0 0 9 ) 847\u2013855 and the pitch circles of the generated gears, the arc length of NJ shown in Fig. 12 is equal to the corresponding distance changed according to the shifted amount. follows: E = 2 + \u2212 E (54) where = inv B \u2212 inv E (55) 4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002752_s00034-010-9198-0-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002752_s00034-010-9198-0-Figure2-1.png",
+        "caption": "Fig. 2 The inverted pendulum system",
+        "texts": [
+            " A stabilizing controller in the form of (6) exists, such that the closed-loop sampled-data system in (16) is asymptotically stable if there exist matrices L > 0, R > 0, Xij , Y ij , and Kj satisfying \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2 \u23a3 11 12 13 14 15 16 17 22 23 24 25 26 27 \u2217 \u2217 33 34 35 36 37 \u2217 \u2217 \u2217 44 45 0 0 \u2217 \u2217 \u2217 \u2217 55 BT 1i BT j1 \u2217 \u2217 \u2217 \u2217 \u2217 \u22123T \u22121R 0 \u2217 \u2217 \u2217 \u2217 \u2217 \u2217 \u22123T \u22121R \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5 \u23a6 < 0 (36)\u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2 \u23a3 H11 H12 H13 H14 H15 H16 H17 \u2217 H22 H23 H24 H25 H26 H27 \u2217 \u2217 H33 H34 H35 H36 H37 \u2217 \u2217 \u2217 H44 H45 0 0 \u2217 \u2217 \u2217 \u2217 H55 BT 1i BT j1 \u2217 \u2217 \u2217 \u2217 \u2217 \u22123T \u22121R 0 \u2217 \u2217 \u2217 \u2217 \u2217 \u2217 \u22123T \u22121R \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5 \u23a6 < 0 where 11 = 1 2 (Ai + Ai)L + 1 2 L(Ai + Ai) T + LCT i CiL + Xij + X T ij + 1 2 (Aj + Aj)L + 1 2 L(Aj + Aj) T + LCT j CjL + Xji + X T ji 12 = 1 2 (Bi + Bi)\u03b2iKi ( I + \u039b(t) ) L + LCT i Di\u03b2iKi ( I + \u039b(t) ) L \u2212 Xij + Y T ij + 1 2 (Bj + Bj)\u03b2jKj ( I + \u039b(t) ) L + LCT j Dj\u03b2j \u00d7 Kj ( I + \u039b(t) ) L \u2212 Xji + Y T ji 13 = 1 2 (Bi + Bi)(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L + LCT i Di(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L + 1 2 (Bj + Bj)(I \u2212 \u03b2j )Kj ( I + \u039b(t) ) L + LCT j Dj (I \u2212 \u03b2j ) \u00d7 Kj ( I + \u039b(t) ) L 14 = \u2212d(t)Xij \u2212 d(t)Xji 15 = 1 2 B1iL + 1 2 Bj1L 16 = L(Ai + Ai) T 17 = L(Aj + Aj) T 22 = LT ( I + \u039b(t) )T K T i \u03b2T i DT i Di\u03b2iKi ( I + \u039b(t) ) L \u2212 Y ij \u2212 Y T ij + LT ( I + \u039b(t) )T K T j \u03b2T j DT j Dj\u03b2jKj ( I + \u039b(t) ) L \u2212 Y ji \u2212 Y T ji 23 = LT ( I + \u039b(t) )T K T i \u03b2T i DT i Di(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L + LT ( I + \u039b(t) )T K T j \u03b2T j DT j Dj (I \u2212 \u03b2j )Kj ( I + \u039b(t) ) L 24 = \u2212d(t)Yij \u2212 d(t)Yji 25 = 0 26 = LLT ( I + \u039b(t) )T K T i \u03b2T i (Bi + Bi) T 27 = LLT ( I + \u039b(t) )T K T j \u03b2T j (Bj + Bj) T 33 = LT ( I + \u039b(t) )T K T i (I \u2212 \u03b2i) T DT i Di(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L + LT ( I + \u039b(t) )T K T j (I \u2212 \u03b2j ) T DT j Dj (I \u2212 \u03b2j )Kj ( I + \u039b(t) ) L 34 = 0 35 = 0 36 = LLT ( I + \u039b(t) )T K T i (I \u2212 \u03b2i) T (Bi + Bi) T 37 = LLT ( I + \u039b(t) )T K T j (I \u2212 \u03b2j ) T (Bj + Bj) T 44 = \u22122d(t)LR\u22121L 45 = 0 55 = \u22122\u03b3 2 H11 = 1 2 (Ai + Ai)L + 1 2 L(Ai + Ai) T + LCT i CiL + Xij + X T ij + 1 2 (Aj + Aj)L + 1 2 L(Aj + Aj) T + LCT j CjL + Xji + X T ji H12 = 1 2 (Bi + Bi)\u03b2iKi ( I + \u039b(t) ) L + LCT i Di\u03b2iKi ( I + \u039b(t) ) L + 1 2 (Bj + Bj)\u03b2jKj ( I + \u039b(t) ) L + LCT j Dj\u03b2jKj ( I + \u039b(t) ) L H13 = 1 2 (Bi + Bi)(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L + LCT i Di(I \u2212 \u03b2i) \u00d7 Ki ( I + \u039b(t) ) L \u2212 Xij + Y T ij + 1 2 (Bj + Bj)(I \u2212 \u03b2j )Kj ( I + \u039b(t) ) L + LCT j Dj (I \u2212 \u03b2j ) \u00d7 Kj ( I + \u039b(t) ) L \u2212 Xji + Y T ji H14 = \u2212( T \u2212 d(t) ) Xij \u2212 ( T \u2212 d(t) ) Xji H15 = 1 2 B1iL + 1 2 Bj1L H16 = L(Ai + Ai) T H17 = L(Aj + Aj) T H22 = LT ( I + \u039b(t) )T K T i \u03b2T i DT i Di\u03b2iKi ( I + \u039b(t) ) L + LT ( I + \u039b(t) )T K T j \u03b2T j DT j Dj\u03b2jKj ( I + \u039b(t) ) L H23 = LT ( I + \u039b(t) )T K T i \u03b2T i DT i Di(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L + LT ( I + \u039b(t) )T K T j \u03b2T j DT j Dj (I \u2212 \u03b2j )Kj ( I + \u039b(t) ) L H24 = 0 H25 = 0 H26 = LLT ( I + \u039b(t) )T K T i \u03b2T i (Bi + Bi) T H27 = LLT ( I + \u039b(t) )T K T j \u03b2T j (Bj + Bj) T H33 = LT ( I + \u039b(t) )T K T i (I \u2212 \u03b2i) T DT i Di(I \u2212 \u03b2i)Ki ( I + \u039b(t) ) L \u2212 Y ij \u2212 Y T ij + LT ( I + \u039b(t) )T K T j (I \u2212 \u03b2j ) T DT j Dj (I \u2212 \u03b2j )Kj ( I + \u039b(t) ) L \u2212 Y ji \u2212 Y T ji H34 = \u2212( T \u2212 d(t) ) Yij \u2212 ( T \u2212 d(t) ) Yji H35 = 0 H36 = LLT ( I + \u039b(t) )T K T i (I \u2212 \u03b2i) T (Bi + Bi) T H37 = LLT ( I + \u039b(t) )T K T j (I \u2212 \u03b2j ) T (Bj + Bj) T H44 = \u22122 ( T \u2212 d(t) ) LR\u22121L H45 = 0 H55 = \u22122\u03b3 2 Then the subsystem controller can be expressed as Ki ( I + \u039b(t) ) L = Ki ( I + \u039b(t) ) LL (37) where L is determined by (4). The proof is omitted for lack of space. 3 Illustrative Example In this section, an example is presented to show the validity of our control scheme. We use our controllers to control an inverted pendulum. Figure 2 shows the inverted pendulum system (or the cart-pole system). The equations of the inverted pendulum system are as follows: x\u03071 = x2 x\u03072 = g sinx1 \u2212 ml cosx1 sinx1 mc+m l( 4 3 \u2212 m cos2 x1 mc+m ) + cosx1 mc+m l( 4 3 \u2212 m cos2 x1 mc+m ) u where g = 9.8 m/s2 is the acceleration due to gravity, mc is the mass of the cart, m is the mass of the pole, l is the half length of the pole, and u is the applied force (control). We chose mc = 1 kg, m = 0.1 kg. Hence, x\u03071 = x2 x\u03072 = \u22120.02x1 \u2212 0.25x3 2 + u We choose the fuzzy membership function as \u03bc1(x1) = 1 \u2212 x2 1 and \u03bc2(x1) = 1 \u2212 \u03bc1(x1)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure8-1.png",
+        "caption": "Fig. 8. Torsion loadcase: inverse force identification",
+        "texts": [
+            " Another option would be to include a support point on the neutral line of the vehicle body in the measurements. This point could then be constrained in Z-direction to avoid the rotational rigid body mode. But also this option would require extra measurement work. However, when using an inverse force identification, there is no need for additional measurements. Instead of forces, displacements are now applied to the front domes: i.e. a fixed displacement (d1 and d2) in Z-direction at both front domes in order to achieve a rotation of 1 degree as shown in Fig. 8. By means of a matrix inversion of the coupled matrix Hc, the forces at the front domes necessary for this displacement and the torsional stiffness in terms of moment (Nm) per degree can then be calculated (see Eq. 3). In figure 9 the resulting forces F1 and F2 are plotted and they typically convert to a constant value around 0 Hz which make it possible to extract the applied torque and therefore the static torsional stiffness. The previous described FBS approach makes it theoretically possible to extract static stiffness values from free-free FRF measurements"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001559_acemp.2007.4510568-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001559_acemp.2007.4510568-Figure9-1.png",
+        "caption": "Fig. 9. Path of the permanent magnets leackage flux",
+        "texts": [
+            " 3D flux paths In hybrid excitation synchronous machines, the flux paths are not only 2D paths but 3D paths exists also. Those paths can be classed in three categories. The first categorie correspond to paths that cross only once the air gap, that is, they pass through only one pole (homopolar paths). They lead to a non-zero mean value of the flux (Fig. 8). The second category is the one of paths that cross the air gap twice, once through each poles (bipolar paths). The third categorie contains all leackge paths. One of them is illustrated in the Fig. 9. It\u2019s a possibel path for the permanent magnets leackage flux. Indeed, a part of the permanent magnets flux doesn\u2019t cross the actif air gaps and stay localised only in the rotor [3]. The final magnetic equivalent circuit of the homopolar hybrid excitation synchronous machine is obtained by using the develloped analytical model for 2D paths and integrating the 3D paths defined above (Fig. 10). Reluctances defined for this model are indicated in the nomenclature. They are calculated such as the ones of the 2D model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002029_j.optlaseng.2010.04.003-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002029_j.optlaseng.2010.04.003-Figure2-1.png",
+        "caption": "Fig. 2. Schematic diagram of the camera stations.",
+        "texts": [],
+        "surrounding_texts": [
+            "The transmission towers have become higher because of the increase of transportation voltage in these years. It makes an estimate of the loading capacity for transmission towers more important. A lot of transmission towers were damaged by the snow disaster in south China, 2008. Most reported studies on loading deformation of transmission towers employed numerical simulation method, which can predict three-dimensional deformation of the structure at low cost, but the predicted result is often far away from the real experiment, due to complexity of the loading process [3\u20137]. Moon drew a conclusion that the simulation results cannot agree with the experimental data well when the loading conditions are complex [1]. The loading capacity of transmission towers cannot be calculated accurately only by numerical simulation. Measuring the field deformation of real-sized model and the loading capacity of transmission towers under different load conditions are necessary. The traditional displacement sensors and resistance strain gauges can measure only one-dimensional deformation in limited measurement range, and it cannot measure the whole field deformation of transmission tower [1\u20135]. Traditional sensors are all contact-type device, install of them is difficult or impossible in many cases, especially when access to the structure is prohibited. In order to solve these kinds of problems, optical measurement has been introduced by more and more researchers. Fraser (2000) [5] pasted some fire-resistant artificial markers made of special ll rights reserved. . material on the transmission tower to measure its displacement during the temperature falling from 11001 to room temperature by optical measurement to obtain the diagram of displacement in each step. Jiang (2007) [7]proposed a new scheme controlled by meshwork, while measuring the cement beams deformation of large bridge with loads to decrease the markers and workload when measuring. Xu Fang et al. used digital photogrammetry to measure the deflection of steel structure. Huang measured windinduced liberation of a high altitude cantilever transmission tower at the scene by photogrammetry to solve the problem that measuring the inaccessible object dynamically [7\u201312,16]. In this paper, a non-contact 3D optical static deformation measurement system called XJTUSD is developed by a Xi\u2019an Jiaotong University in China, in order to monitor the 3D deformation of realsized transmission tower during loading test. The key technologies of close range industrial photogrammetry are studied, including the disposition of markers, the disposition of camera stations, deformation of single step, measurement coordinates alignment of different steps, the registration method for corresponding points in multistep, calculation and display for the deformation. The deformation of the whole transmission tower with different loads is obtained which is helpful for further quantitative analysis."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure5.2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure5.2-1.png",
+        "caption": "Figure 5.2 Flapper valve example of a nonlinear function",
+        "texts": [
+            " There is one additional form of nonlinearity that must be mentioned here and that is the pure time delay also referred to as a transport delay or dead time. This type of nonlinearity falls into neither of the above two categories and will therefore be discussed as a separate topic. A good example of this form of nonlinearity is found in hydraulics and pneumatics where the flow equations involve a number of nonlinear functions. If we consider a simple flapper valve found in typical hydraulic servos (see Figure 5.2) we have a nonlinear relationship between the flapper displacement, and the flow through the nozzle. 160 Dealing with Nonlinearities This type of equation does not readily fit into the linear transfer function methodology presented so far in this book, i.e. Q= KVxV \u221a Pi\u2212Po where Q is the flow through the flapper valve, KV is a valve flow/ geometry coefficient, xV is the flapper valve displacement, and Pi\u2212Po is the valve pressure drop. We have ignored here the flapper travel limits for the purpose of this exercise since this would fall into the category of discontinuous nonlinearities to be discussed in the next section"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002063_robio.2010.5723474-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002063_robio.2010.5723474-Figure3-1.png",
+        "caption": "Fig. 3. Chart for solving q4",
+        "texts": [
+            " On the following, an obstacle avoidance based motion planning algorithm will be deduced. To the 7DOF redundant manipulator discussed in this paper, for a specific Destination Matrix H0 7 of it, the origin of coordinate system \u03a36 can be verified solely. Because the joint angle value q7 just change the orientation relationship of coordinate system \u03a36 and coordinate system \u03a37 while it have no affect on the position relationship of the two coordinates systems. Then the origin of the coordinate system \u03a36 can be calculated (Fig.3.): p0 6 = H0 7 \u239b \u239c\u239c\u239d 0 0 \u2212d7 0 \u239e \u239f\u239f\u23a0 = \u239b \u239c\u239c\u239d p6x p6y p6z 0 \u239e \u239f\u239f\u23a0 (3) The same case, the joint angle q1 will not affect the position relationship of coordinate system \u03a32 and coordinate system \u03a30 which means that the origin of coordinate system \u03a32 is p0 2 = ( 0 0 d1 )T . Based on the analysis above, it is obvious that when the Destination Matrix is fixed, the position of p0 6 and p0 2 are also verified. Then the distance of these two points can be calculated: L26 = \u2225\u2225p0 6 \u2212 p0 2 \u2225\u2225 (4) Based on the physic structure of the manipulator, it is easy to find that joint 2, joint 4 and joint 6 form a triangle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002514_jfm.2011.256-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002514_jfm.2011.256-Figure1-1.png",
+        "caption": "FIGURE 1. (Colour online available at journals.cambridge.org/flm) Principle of the float glass process: molten glass floats on top of molten tin. While it cools down, the glass is stretched by top rollers and end rollers to achieve the desired thickness. The float part is about 50 m long and 10 m wide.",
+        "texts": [
+            " As a consequence, buckling instabilities can also be expected, yielding a mechanism for morphogenesis, as for organogenesis in plants (Steele 2000) or fingerprint formation (Ku\u0308cken & Newell 2004), or convolutions of the brain (Toro & Burnod 2005). We note, however, that, except in this last study, models were built upon purely elastic approaches. In the glass industry, thin sheets are given prescribed shapes and thickness by applying mechanical forcing while they cool down (Pearson 1985). In the second part of this paper we will consider the case study of a floating sheet, which is motivated by both plate tectonics and glass industry. The glass float process (Pilkington 1969) is used for a continuous production of thin glass (figure 1). At one end of a long bath of liquid tin, molten glass is poured from the furnace and spreads out forming a layer of floating glass (tin is denser than glass), cooling down as it gets farther from the furnace. In order to make the layer thinner than the capillary length, a set of rollers are used to stretch the floating sheet, which comes out solidified at the other end of the bath, after it has undergone a transverse compression. The corresponding compressive stress may induce buckling. Taylor (1968) demonstrated the buckling of viscous threads and sheets in various experimental geometries"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001167_s00158-009-0398-9-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001167_s00158-009-0398-9-Figure1-1.png",
+        "caption": "Fig. 1 Schematic of the actuation force modeled using 6 force design variables in a cycle. The force value between two design variables is computed using cubic splines",
+        "texts": [
+            " The displacement amplitude is selected to represent lift in an actual flapping-wing MAV, although this relation may not be truly one-to-one. A direct lift constraint is considered in Section 6.2 for the hovering insect flight application. To solve the two-objective problem, the work is minimized for a specified target of displacement amplitude \u0393Ti (Kurdi and Beran 2008a): min b W (1a) subject to max(|xca|) \u2265 \u0393Ti i = 1, ..., N, (1b) where b is the vector of design variables, which consist of the magnitudes of the actuation force f at equally spaced locations in one cycle (see Fig. 1), and the circular forcing frequency, \u03c9. The time-periodic response in the direction of actuation force f is denoted by xca. The subscripts c and a refer to the cyclic response and degree-of-freedom in direction of actuation, respectively. To construct the tradeoff curve of minimum work and maximum amplitude the optimization problem is repeatedly solved for a total number of N times. In each time, the target displacement \u0393Ti is increased and the minimum work is found. For each design i on the tradeoff curve, the optimal actuation force (Fig. 1) and frequency are computed. The tradeoff curve, although expensive to compute, provides the designer with all possible optimal designs of the two objectives, thereby allowing selection of the best design fitting user criteria. Alternatively, the optimization problem of minimum work and maximum amplitude under limited force may be formulated by maximizing the response amplitude for a set of constraints on the work objective. Although this should lead to similar results to that of (1) the latter formulation is more effective for the design of nonlinear systems due to dependence on initial conditions (more on this in Section 5.2.1). The force design variables (6 variables in Fig. 1) are allowed to change in the interval [\u2212 fmax, fmax ] during the optimization search. With no additional constraints on the difference in magnitude between two adjacent variables in time, this may result in a force trajectory infeasible using realistic actuators. This difficulty may be avoided by enforcing the limitation of the actuator once it is known. This is applied for the MAV application; see Section 6. The period T of the response cycle is equal to the forcing period T = 2\u03c0/\u03c9. The time interval t is scaled by T leading to a scaled cycle s of length 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001854_gt2009-59580-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001854_gt2009-59580-Figure1-1.png",
+        "caption": "Figure 1. Blade-Disk Interface Zone",
+        "texts": [
+            " The disk stiffness and mass matrices in the coordinates ( )dd Yq , are then \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 = YY CBd qq CBd CBd K K K , , , 0 0 (10) and \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 = YY CBd Yq CBd qY CBd qq CBd CBd MM MM M ,, ,, , . (11) To complete the blade-interface-disk modeling, it remains to relate the response of the interface degrees-of-freedom on the disk side to those on the blades. In this regard, two situations can be envisioned. In the first, the interface degrees-of-freedom of the disk are not collocated with those of the blades, see Fig. 1a, and a physical interface zone has been created. Assuming that there are no external forces acting on the interface zone of blade b, the disk-side and blade-side interfaces, bdF , and bF can be expressed as \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 + \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 = \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 + \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 =\u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 b bd bd b bd bd b bd bbdbb bdbdbdb b bd bbdbb bdbdbdb b bd Y Y K Y Y M Y Y KK KK Y Y MM MM F F , , , , , ,, ,,, ,, ,,, && && && && (12) where dbdb M , , bdb M , , dbb M , , etc. are the mass and stiffness matrices of the finite element model of the contact zone. Note that bb M , and bb K , are in general not equal to II physb M , and II physb K , , although they 2 Copyright \u00a9 2009 by ASME ms of Use: http://www.asme.org/about-asme/terms-of-use Dow If a detailed finite element model of the interface zone has not been carried out, e.g. see Fig. 1b, it is still possible to introduce relative motions at the blade-disk interface by adopting Eq. (12) but with chosen mass and stiffness matrices dbdb M , , bdb M , , dbb M , , etc. For example, assuming that there is only flexibility at the interface (no inertial effect), one can select 0 ,,,, ==== dbbbdbbbdbdb MMMM (13) and KKKKK dbbbdbbbdbdb \u02c6 ,,,, =\u2212=\u2212== (14). Without further information, it is convenient to select the matrix K\u0302 in terms of the stiffness matrices of the interface degrees-of-freedom on the blade ( II physb K , ) and disk ( II physdb K , ) sides as \u239f \u23a0 \u239e \u239c \u239d \u239b += II physdb II physb KKkK ,, \u02c6 (15) where the scalar k is an adjustable, dimensionless parameter that describes the stiffness of the blade-disk interface"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001883_icndc.2010.72-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001883_icndc.2010.72-Figure2-1.png",
+        "caption": "Figure 2. Sequence of gait phases in one walking cycle",
+        "texts": [
+            " Inevitably, such strong impact force will affect the stability of the robot. This section presents compliance control to compensate impact/contact forces generated between the landing foot and the ground. In the study, we assume the foot remains horizontal in every moment during walking. The humanoid robot\u2019s gait phase consists of the singlesupport phase and the double-support phase in one cycle during walking. If only one foot is on the ground, the biped robot is in the single-support phase. On the contrary, the robot is in the double-support phase. Fig. 2 shows the gait phase in one walking cycle. When the single-support phase is changed to double-support phase, the robot\u2019s swing foot starts to contact with the ground surface. At this moment, the robot gets a big force that can make the robot unstable. The biped robot uses the inverted pendulum model to plan the dynamic walking gait. In paper [10], a realtime pattern generator is proposed using three-dimensional inverted pendulum. When generating the robot\u2019s walking trajectory, it is usually expected the swing leg reaches the ground at the end of the single-support phase and leaves at the start of double-support phase"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000882_4243_2008_022-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000882_4243_2008_022-Figure2-1.png",
+        "caption": "Fig. 2 A disposable sensor for blood gas analysis containing an extracorporeal loop, from [17]. The three sensor spots (for pH, O2, and CO2) are interrogated via three 200-\u00b5m fiber optic cables",
+        "texts": [
+            " Moreover, various colored species may be contained in a matrix with optical properties similar to the species of interest so that they cannot be recognized by direct spectroscopy. This situation gave rise to the development of fluorosensors, in which the analytical information is mediated by some sort of indicator chemistry, usually deposited in the form of a thin sensor film. The film is composed of an analyte-permeable polymer that contains the chemically responsive probe. This film can be used in various ways but mostly in the form of a sensor spot as shown in Fig. 2 which is a schematic of the sensing unit of a widely used medical system. Such films may, however, also be deposited inside a reaction bottle, a microwell plate, or at the core of a fiberoptic waveguide as shown in Fig. 3 where a fluorogenic enzyme substrate in a polymeric solid support is placed on either the core or the distal end of the fiber. The support also may contain indicator probes for analytes such as pH, oxygen, and the like. Other examples of such \u201copt(r)odes\u201d are listed in Table 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000931_bfb0119383-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000931_bfb0119383-Figure4-1.png",
+        "caption": "Figure 4. (Left) A picture of the paper lifting robot. (Pdght) Details of the sticky foot mechanism of the robot Fiat.",
+        "texts": [
+            " Our conclusions were tha t unwanted slip is bad enough to require additional sensors, bu t not bad enough to prevent effective manipula t ion and locomotion. We also discovered tha t we need a be t te r approach to control, coupling the wheel servos and a t tending to the nonholonomic na ture of the device. The Mobipulator can move single sheets of paper in the x and y directions when the papers rest directly on a desktop. To cope with areas of desktops tha t are crowded with stacks of paper we designed the mobile robot called Fiat (see Figure 4). In the future, we plan to combine the functionali ty of F ia t and the Mobipulator into a single mobile robot. Fiat can lift a single piece of paper from a stack and move it to a different location on the desktop. The mechanical design for F ia t was mot iva ted by the goal of moving paper in the vertical direction, in order to stack sheets one above the other deterministically. The manipulator of this robot is a foot made sticky with a supply of removable tape. A spool of removable adhesive tape provides a supply of \"fresh stickiness\" for the foot (see Figure 4). The manipulator needs a way to attach the foot to a piece of paper and to release the paper from the foot; this is accomplished by integrating the manipulator with the locomotive design. The foot travels on an eccentric path relative to the free rear wheels. In one direction, the foot emerges below the wheels, using the weight of the robot to stick the foot to the paper. As the eccentric motion continues, the foot moves behind the wheels, and lifts up the paper. The opposite motion detaches the foot from the paper, using the rear wheels to hold down the paper while the sticky foot is freed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003934_icinfa.2013.6720491-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003934_icinfa.2013.6720491-Figure2-1.png",
+        "caption": "Fig. 2. Free body diagram of the quadrotor.",
+        "texts": [
+            " An effective solution is to obtain the control gains in simulation by employing the derived mathematical model of the quadrotor and then use these gains as references to finely tune the controller in actual flight tests. Although these obtained gains are not accurate enough for real system operations, it already narrows down the range of searching for the optimal control gain values. The nonlinear dynamic equations of the quadrotor are derived from first principles. The parameters of the dynamic model are obtained by building a physically accurate model of the quadrotor in SolidWorks. The 3D drawing can be seen in Fig. 1. As shown in Fig. 2, two reference frames are used to represent the state vectors including a North-East-Down (NED) coordinate system regarded as an inertial reference frame and a body-fixed coordinate system. NED inertial frame complying with the right-hand rule is denoted by {\ud835\udc4a} = {\ud835\udc7e\ud835\udc99,\ud835\udc7e\ud835\udc9a,\ud835\udc7e\ud835\udc9b}, where the unit vector along \ud835\udc7e\ud835\udc9b is aligned with the direction of gravity. The origin of bodyfixed frame is attached to the center of mass of the quadrotor. The frame is denoted by {\ud835\udc35} = {\ud835\udc69\ud835\udc99,\ud835\udc69\ud835\udc9a,\ud835\udc69\ud835\udc9b} with the unit vectors \ud835\udc69\ud835\udc99, \ud835\udc69\ud835\udc9a and \ud835\udc69\ud835\udc9b pointing to forward, rightward and downward directions relative to the fuselage, respectively",
+            ") of the quadrotor with respect to the \ud835\udc7e\ud835\udc99, \ud835\udc7e\ud835\udc9a and \ud835\udc7e\ud835\udc9b axes, respectively, and the linear velocity {\ud835\udc62, \ud835\udc63, \ud835\udc64} relative to the inertial reference frame. The nonlinear dynamics of the aircraft for both translational and rotational can be expressed in Newton-Euler formalism: { \ud835\udc64\ud835\udc39 = \ud835\udc5a \u22c5 \ud835\udc5f \ud835\udc40 = \ud835\udc3c \u22c5 \ud835\udc4f?\u0307? + \ud835\udc4f\ud835\udf14 \u00d7 (\ud835\udc3c \u22c5 \ud835\udc4f\ud835\udf14) (1) where \ud835\udc64\ud835\udc39 is the total force acting on the center of mass in {\ud835\udc4a}, \ud835\udc5a denotes the mass of the vehicle, \ud835\udc5f is the position vector from the origin of the NED coordinate to the c.g., \ud835\udc40 denotes the total moment acting about c.g., \ud835\udc3c is moment of inertial and \ud835\udc4f\ud835\udf14 denotes the angular velocity in {\ud835\udc35}. As described in Fig. 2, rotor 1 is located on positive \ud835\udc69\ud835\udc99 direction coinciding with the direction of forward flight. Rotor 2 is on positive \ud835\udc69\ud835\udc9a , 3 on negative \ud835\udc69\ud835\udc99 and 4 on negative \ud835\udc69\ud835\udc9a . The total forces acting on the quadrotor are the gravity in \ud835\udc7e\ud835\udc9b direction and the thrust from each rotor \ud835\udc47\ud835\udc56 perpendicular to the rotor plane along \u2212\ud835\udc7e\ud835\udc9b direction. Each spinning rotor yields an opposing aerodynamic moment in {B}, denoted by \ud835\udc40\ud835\udc56. The directions of \ud835\udc401 and \ud835\udc403 are perpendicular to the rotor plane along \u2212\ud835\udc69\ud835\udc9b while \ud835\udc402 and \ud835\udc404 act in \ud835\udc69\ud835\udc9b direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002907_ol.35.000456-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002907_ol.35.000456-Figure1-1.png",
+        "caption": "Fig. 1. Coordinate system of optical components at a general TNLC cell.",
+        "texts": [
+            " We can obtain the zero transmittance at a certain wavelength by varying the angles of polarizers and then by solving appropriate equations related to the intensity and polarization of light, the cell parameters can be determined. In order to determine easily the LC parameters of the wedge-type TNLC cell in an ordinary laboratory, we used a rotating polarizer method with single-wave source. This is easy to use and of low cost. Furthermore, it can provide precise results for the retardation and the twist angle of LC cell. The optical transmission of a twisted nematic LC cell with the optical configuration of Fig. 1 is given by [17] T = cos cos r \u2212 p + r sin sin r \u2212 p 2 + 2 2 sin2 cos2 r + 2 0 \u2212 p , 4 where p is the angle between the transmission axes of input and output polarizers, 0 is the angle between transmission axis of input polarizer and input LC director, and = r 2 + 2 1/2, = d n/ , n = ne 1 + z sin2 \u2212 no, z = ne/no 2 \u2212 1. Here, is the pretilt angle, is the wavelength of light source, and ne and no are the extraordinary and ordinary refractive indices of the LC material, respectively. From Eq. (4), it can be seen easily that for a null condition of T, if sin =0, T = cos2 r \u2212 p = 0, namely, r \u2212 p = 2 \u00b1 n "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000667_s00170-007-0927-x-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000667_s00170-007-0927-x-Figure3-1.png",
+        "caption": "Fig. 3 Section view of the test rig",
+        "texts": [
+            " The torsional force is transmitted to the measurement beams through the pull lever on both sides. And by measuring the deformation of the beam through strain gauges, the frictional force is determined .The pin and the rolling ball bearings are used to facilitate the mounting of the beam along with the bearing plate at any required crank angle. The test was first carried out with commercial-based oil. Then the base oil containing an oil fortifier was sent and the friction force was measured at each 15\u00b0 crank angle up to 360\u00b0 [3]. 2.1.2.1 Construction and working principle Figure 3 shows the sectional view of the test rig which was designed and manufactured to measure the oil film thickness by considering the frictional force [4]. A commercial-based oil and oil fortifier with the concentration volume ratio of 5% added to the base oil were used for the experiment, lubricant characteristics are given in Table 1. A commercially available Babbitt alloy bearing was used for the test bearing. The test bearing had a central circular groove. Four oil supply holes were placed in the groove"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure2-1.png",
+        "caption": "Fig. 2 Profile models on which side relieving and crowning is applied: (a) side relieving, (b) crowning",
+        "texts": [
+            " Premature contact at the tips or excessive contact pressures at the end of the teeth give rise to noise and/or gear failures. In order to reduce these causes of excessive tooth load, profile modification is a usual practice [19]. One of these modifications is tooth tip relief. Here, in order to maintain fine gear teeth meshing, a small amount of material is removed from the peak area of the gear teeth. Tooth tip relief is classified into two parts: addendum and dedendum relief (Fig. 1). When comparing the two relief types, addendum relief seems to be widely used. Figure 2 shows Crowning and Side Relief, another type of modification. These two modifications are applied along the axis of the gear teeth surfaces (Fig. 2). Crowning involves removal of a small amount of material from the gear teeth layer, starting from the center towards the edge of the teeth and causing the teeth surfaces to acquire convex-like shapes. Crowning helps to maintain teeth contact in the middle part of the gear teeth as well as preventing edge contact resulting from lower load carrying capacity. In addition, gear assembly is usually character- ized by some misalignments. If good tolerances are assigned to the gears, the effects of these misalignments are not easily felt"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000815_mcs.2008.927961-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000815_mcs.2008.927961-Figure1-1.png",
+        "caption": "FIGURE 1 An eight-degree-of-freedom autonomous agricultural navigation platform. Each wheel is equipped with a driving motor and a steering motor. Thus, each wheel can be individually steered and driven.",
+        "texts": [
+            "00\u00a92008IEEE Automatic navigation of agricultural vehicles, such as tractors and harvesters, can increase the level of automation in the agricultural process, thereby lessening the human workload [1], [2]. In the agricultural environment, where vehicles move on uneven terrain, achieving maneuverability and mobility can be difficult, particularly when the path curvature is small or vehicle orientation and path tracking must be independently controlled. To obtain the desired maneuverability and mobility, one approach is to individually actuate the drive and steering functions of each wheel on a four-wheeled vehicle. Figure 1 shows a prototype system for agricultural purposes. Figure 2 shows a vehicle schematic and block diagram model for closed-loop control of the steering and driving motors. For sensing and navigation, this prototype vehicle uses a combination of machine vision, a differential global positioning system (DGPS), and an inertial measurement unit (IMU). The focus of this work, however, is on path-following control. Although the vehicle is highly maneuverable due to its eight independent actuators, precise control is difficult due to the overconstrained nature of the actuation",
+            " from Shanghai Jiaotong University, China, in 2004. He is currently a Ph.D. candidate in the Department of Mechanical Engineering. His research interests include computer vision, motion control, and robotics. Jun Zhou received the Ph.D. from NanJing Agricultural University in 2003. He is currently a professor in the department of agricultural machinery in NanJing Agricultural University. His research interests include image processing and computer vision. A n arm-wrestling robot called Robo Armwrestler (Figure 1) has recently been developed in our Intelligent Control and Robotics Laboratory to benefit the health care of senior citizens. The motivation for this project, which is supported by the Korean Government, is to reduce social welfare costs and to improve the quality of life of the elderly population by meeting their physical and mental needs. In recent decades, Korea\u2019s aging population has increased by 35%, well over the standard 30% for an aging society. Our vision is to realize humanoid robots that have entertaining functions such as arm wrestling and chess playing, as well as service functions such as errands",
+            " Another effort relating to a humanoid robot arm has been in the field of prosthetic devices, such as the Utah Artificial Arm [3]. However, it does not appear that prosthetic devices are suitable for arm wrestling, in which strong arm force is required. Several practical arm-wrestling devices have been patented as amusement units or as units for developing and strengthening wrist and arm muscles [4]\u2013[7]. These devices are classified roughly into three types according toDigital Object Identifier 10.1109/MCS.2008.927963 FIGURE 1 The arm-wrestling robot Robo Armwrestler. This robot, which is intended for use in elderly health care, was developed in the Intelligent Control and Robotics Laboratory at Konkuk University in Korea. 1066-033X/08/$25.00\u00a92008IEEE"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002212_lars.2010.27-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002212_lars.2010.27-Figure2-1.png",
+        "caption": "Fig. 2. Traction and resistive forces on the mobile robot.",
+        "texts": [
+            " (4) Replacing equation (4) on (3) and writing in the Pfaffian\u2019s form, we have [ \u2212 sin\u03c8 cos\u03c8 \u2212x\u2032 CIR ] \u23a1\u23a3 x\u0307 y\u0307 \u03c8\u0307 \u23a4 \u23a6 = A(q)q\u0307 = 0 From Fig. 1, the velocities q\u0307 can be represented as q\u0307 = S(q)\u03b7, (5) where S(q) = \u23a1 \u23a3 cos\u03c8 \u2212x\u2032 CIR sin\u03c8 sin\u03c8 x \u2032 CIR cos\u03c8 0 1 \u23a4 \u23a6 , \u03b7 = [ \u03c5x\u2032 \u03c9 ] . S(q) is a full rank matrix, whose columns are in the null space of A(q), i.e. ST (q)AT (q) = 0 It is interesting to note that because dim(\u03b7) = 2 < dim(q) = 3, the equation (5) describes the kinematic of a sub-actuated robot with the nonholonomic constraint given by (3). The forces and moments acting on the robot are presented in Fig. 2, where Fx\u2032 i and Rx\u2032 i are the traction forces and the longitudinal resistive forces acting on each wheel. The Pioneer 3-AT has two motors, each one for driving the wheels of one side of the robot. So we can assume that Fx\u20321 = Fx\u20323 and Fx\u20322 = Fx\u20324, which reduces the longitudinal skid. The lateral forces Ry\u2032 i on each wheel are consequence of the lateral skid. The resistive momentMr around the center of mass, which opposes to the moment M , is induced by the forces Ry\u2032 i and Rx\u2032 i. The motion equations w"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001857_10402000903283292-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001857_10402000903283292-Figure1-1.png",
+        "caption": "Fig. 1\u2014Normal stresses in the grease generating a lift to separate the seal from the shaft.",
+        "texts": [
+            " No exact values for the normal stress could be measured on the Weissenberg rheogoniometer because of the dependence on the yield stress and zeroing problems of the normal force. Therefore, Binding, et al. (11) continued the work on a torsional balance rheometer, and here the normal stress was balanced by an external load, to determine absolute values of the normal stress. They showed that the normal stress in grease is of the same order of magnitude as the shear stress. This normal stress effect in grease and the relatively low contact pressures between the seal lip and the shaft might result in a significant contribution to film formation. Figure 1 shows how 308 D ow nl oa de d by [ G eo rg ia T ec h L ib ra ry ] at 1 0: 05 1 2 N ov em be r 20 14 a = vicinity width air side b = contact width c = position minimum film height d = vicinity width lubricant side Finertia = inertial force Flift = specific lift force Flip = specific lip force Frheology = specific normal stress force F\u00b5EHL = specific micro-elastohydrodynamic force h1 = film height at edge contact h2 = film height at c/b = 1/3 hg = grease height ho = minimum film height h(x) = film thickness function K = constant Kc = constant Mfriction = frictional torque m = power law index N1 = first normal stress difference N2 = second normal stress difference n = power law index nc = power law index R = rim radius r = radius T = temperature u = shaft surface speed x1 = integration boundary x2 = integration boundary \u03b1 = lip air side angle \u03b2 = lip lubricant side angle \u03b3\u0307 = shear rate \u03b3\u0307m = measured shear rate \u03b5 = gap error \u03b7 = viscosity \u03b7m = measured viscosity \u03b7o = low shear rate viscosity \u03b7r = real viscosity \u03b7\u221e = high shear rate viscosity \u03b7\u221e = base oil viscosity \u03bb1 = relaxation time constant \u03c1 = grease density \u03c3zy = shear stress \u03c4\u03b3 = yield stress \u03d5 = lip angle \u03c81 = normal stress constant = angular velocity the normal stress effect may load the seal lip and generate a lift force",
+            " The Weissenberg rod-climbing effect has also been observed for grease by the authors but is rather small. This is expected to be due to the yield stress behavior. The first normal stress difference N1 acts in a direction that is orthogonal to the plane of shear. The second normal stress difference N2 acts in a direction that is parallel to the plane of shear and perpendicular to the direction of shear. N2 is negative, normally much smaller than N1, and therefore in many viscoelastic fluids negligible. In the radial lip seal contact, N1 will tend to lift the seal from the shaft (see Fig. 1) and enhance lubricant film formation. N2 will force lubricant to move from the axial direction into the direction of the highest shear rate and could be a source for lubricant supply into the contact and could also play a role in the pumping action of the seal. N2 is assumed to be small compared to N1 and is neglected in the present study. In classic constitutive models, a relation between shear rate and the first normal stress is available (Barnes, et al. (6)). This relation follows a power law behavior over a range of shear rates and reads N1 = \u03c81\u03b3\u0307 m [5] In classic models like the second-order Maxwell model or the Oldroyd-B model, the constant m = 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003086_pime_conf_1966_181_311_02-Figure8.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003086_pime_conf_1966_181_311_02-Figure8.4-1.png",
+        "caption": "Fig. 8.4. Typical pressure distribution",
+        "texts": [
+            " The influence of the inlet position has also been examined, and in practice this takes account of the depth of submersion or availability of lubricant for the cylinders. The results are presented in terms of side-leakage factors by forming the ratio of the force component on the finite cylinder to the force component on a section of an infinite cylinder of equal axial width. These factors can then be used directly to modify the Martin theory predictions summarized by equations (8.7). A typical pressure distribution for a value of R/ho of lo3 and I/R equal to 0.025 is shown in Fig. 8.4. The diagram shows that the important pressure generating zone is restricted to a region about 0.1R in length, even though the build-up of pressure is permitted to occur from an inlet section at a distance R from the centre-line. The axial decay of pressure towards the edge of the cylinder is shown, and it will be noted that the cavitated region is separated from the load-carrying region by a curved boundary. The pressure distribution shown in Fig. 8.4 is typical of all the solutions obtained during this investigation. Side-leakage factors for the normal force components P, and P, are shown for values of R/ho ranging from loa to lo6 in Figs 8.5 and 8.6. The results cover cylinders of half length-radius ratio up to unity, and it can be seen that greater lengths are unlikely to produce large changes in the side-leakage factor until the cylinders become very long. The angle which the resultant normal force on the VoI I81 Pt 3 0 at WEST VIRGINA UNIV on June 5, 2016pcp"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure11-1.png",
+        "caption": "FIGURE 11. Combination of four powers to move a weight of 1000 talents with a force of 5 talents (Mechanics 2.29). Drawing of a figure in Ms L (Drachmann, The Mechanical Technology, p. 90).",
+        "texts": [
+            "21: These things having been done, if we imagine the chest AB placed on high, and we tie the weight to the axle EZ, and the pulling force to the wheel X , neither side will go downwards, even if the axles are turning easily and the engagement of the wheels is fitted nicely, but the force will balance ( , o o \u0301 ) the weight as in a balance ( , \u00b4 o\u0302 o\u0301 ). But if we add a little more weight to one of them, the side where the weight is added will sink ( \u0301 ) and go downwards, so that, if just the weight of one mina is added to the force of five talents, it will overpower ( \u0301 ) the weight and pull it.45 As a final tour de force Heron describes how all the powers except for the wedge can be combined to achieve the same mechanical advantage of 200:1 (2.29; Fig. 11). The discussion of the powers in combination prompts Heron to remark on a further important aspect of their operation, the phenomenon of delay 45 Heron, Dioptra, 37 (Opera, vol. III, pp. 310.20\u2013312.2); Engl. trans. Drachmann, The Mechanical Technology, p. 26, slightly modified. Throughout the passage \u201cforce\u201d translates \u0301 and \u201cweight\u201d \u0301 o . Pappus, Pappi Alexandrini collectionis quae supersunt, VIII, vol. III, p. 1066.19\u201331, expresses the same idea in the same language. See also Heron, Mechanics, 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001877_1.3084237-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001877_1.3084237-Figure8-1.png",
+        "caption": "Fig. 8 With f of 0.25, von Mises stress over the central cross section of \u201ea\u2026 the coated half-space with coating of 3 m thick, \u201eb\u2026 the half-space made of the coating material, and \u201ec\u2026 the coated half-space as in \u201ea\u2026, but with larger depth",
+        "texts": [
+            " 1 are also pplied here for consistency. For this problem, a Hertzian contact idth for two cylinders with Young\u2019s moduli of 210 GPa is used s reference for dimensionless size quantities: a=165.4 m. Figre 8 shows the von Mises stress distribution a for the coated ylinder and b for the uncoated one made of coating material. he studied depth at the central cross section is 10 m in order to ompare the stress field. A contour plot for the coated cylinder in larger depth with the same format as in Figs. 7 b and 7 c is hown in Fig. 8 c for a direct comparison. The mismatch of odulus is attributed to the obvious discontinuity of von Mises tress at the interface in Fig. 8 a . The softer substrate of the oated cylinder is subject to a lower stress but causes higher stress n the coating. Conclusions Based on the emerging need to analyze the contact performance f periodic surfaces in line contacts, this paper proposes two new ethods, i.e., DC-FFTS and DC-CC-FFT. These two methods, long with various others, are used to solve the SPLC problems, nd the comparison is illustrated and discussed. The DC-FFTS is a odified DC-FFT method with a summation of ICs from several eriods, and this concept of using summation of ICs can also be pplied to multigrid methods"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003147_1.4007806-Figure16-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003147_1.4007806-Figure16-1.png",
+        "caption": "Fig. 16 (a) Bearing with distinct multiple faults, (b) bearing with an axial fault, and (c) experimental equipment",
+        "texts": [
+            "5, the rate of the ADFH reduction will differ with the HAC when the second defect is added to the bearing, which leads to the MSE values increasing, as shown in Table 1. Since the simulation study proved the two criteria, this method can be examined with the experimental study, which is discussed in the next section. 4.1 Experimental Setup. In order to show the effect of multiple defects in different angles, two types of experiment are considered: (1) Only one row of balls strike the faults. In this experiment, two defects with widths of 1.3 mm and 1.5 mm are created on the outer race. The angle between two defects is 53 deg, as shown in Fig. 16(a). (2) The balls in both rows strike the defect. In this experiment, a defect in the axial direction with a size of 0.6 mm is created. The bearing has 32 balls in two rows (16 balls on each row). The angle difference between two successive balls in one row is 360=16 \u00bc 22:5 deg. Since the balls in another row are on the mid distance of the balls in the other row, the angle between two balls in a different row becomes 22:5=2 \u00bc 11:25 deg, as shown in Fig. 16(b). Therefore, the angle between two successive impacts is 11:25 deg. This experiment is designed to show the effect of crack growth in the axial direction on the ADFH. In order to investigate the proposed method for the diagnosis of multiple defects, a 1207 EKTN double row self-aligning ball bearing is selected. The shaft rotates at a speed of 1620 rpm. The 40 N radial load is applied to the bearing. The defect characteristic frequency of the outer race is 186 Hz. Therefore, the 2nd and 3rd harmonic frequencies are 372 Hz and 558 Hz, respectively. The cage frequency is 11.6 Hz. The second defect is created after the signal related to the first defect was collected. An IMI Sensor type 608A111 accelerometer with a sensitivity of 100 mV=g and a sampling frequency of 10 kHz is mounted on the top of the housing. A DAQ 9233 card is used to digitize the vibration signal. The data is analyzed by LabVIEW programming. The test stand is shown in Fig. 16(c). The procedure of the HFRT is the same as previously described in Sec. 3, however, instead of the bandpass filter, a highpass filter with a cutoff frequency of 3500 Hz is used and the cutoff frequency for the lowpass filter is also adjusted to 700 Hz because three defect harmonics can be extracted from the spectrum. The order of the filters is changed from 1 to 10 and its corresponding time constant is shown in Fig. 17. 4.2 Experimental Data Analysis 4.2.1 Single Fault. The real signal in the time domain and its power spectral density for a single fault are shown in Figs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000558_1.2775516-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000558_1.2775516-Figure2-1.png",
+        "caption": "Fig. 2 Cut view of brush seal test rig",
+        "texts": [
+            " The predictions for damping rely on the appropriate physical characterization of the energy dissipation in the bristle bed and modeled as a hysteretic structural damping type. In the theoretical study, the hysteretic loss factor varies over a certain range since no experimental data were readily available. This paper describes a test rig to perform dynamic load experiments in brush seals and presents a simple identification method to obtain the seal structural stiffness and damping characteristics. The test data, needed to validate predictive models of brush seal performance, bring forward this novel seal technology. Test Rig Description Figure 2 shows a cross section view of the test rig. A long and slender steel shaft 12.7 mm in diameter and an aluminum disk mounted at the shaft end are located inside a cylindrical, thick wall, steel vessel. The disk diameter and thickness equal 163 mm and 25.4 mm, respectively. One end of the shaft is affixed into the bottom of the vessel with two rolling element bearings. The test brush seal is secured at the top of the vessel with an interference fit to the disk. Thus, the simple test system comprises of a cantilever beam whose free end carries a large inertia disk and the test seal element, which offers stiffness and damping connections to ground"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003080_jmer.9000033-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003080_jmer.9000033-Figure3-1.png",
+        "caption": "Figure 3. Connecting rod acceleration.",
+        "texts": [
+            "cos( 21 QQv P \u2212\u2212= (10) And for piston acceleration: \u03b1\u03b8\u03c9\u03b8\u03b1\u03b8\u03c9\u03b8 )2sin(2)2cos(4)sin()cos( 2 2 21 2 1 QQQQaP \u2212\u2212\u2212\u2212= (11) where \u03b1 is the crankshaft rotational acceleration. From Equation 11 and Taylor series, for other parts, can reached the following result: Ranjbarkohan et al. 87 ( ))2cos()cos( 54 \u03b8\u03b8\u03c9\u03bb QQn \u2212= (12) ( ) )sin()2cos()cos( 2 54 \u03b8\u03c9\u03b8\u03b8\u03b1\u03b7 nQQn \u2212\u2212= ( ) )2sin()cos(2)2cos( 2 554 \u03b8\u03b8\u03c9\u03b8 nQQQ +\u2212 (13) That is, \u03bb is the connecting rod rotational velocity and \u03b7 is the connecting rod rotational acceleration. Now the velocity and acceleration of connecting rod\u2019s C.G (Center of gravity) could be calculated. Connecting rod acceleration could be calculated from Figure 3: pgpg aaa / rrr += (14) where pg a / is acceleration vector of connecting rod\u2019s C.G relative the piston, as follow: )( /// pgpgpg rra rrrrrr \u00d7\u00d7+\u00d7= \u03bb\u03bb\u03b7 (15) where rg/p is the displacement vector of connecting rod\u2019s C.G relative the piston, that (Meriam and Kraige, 1998): jisr pg ))sin()cos((/ \u03b2\u03b2 =+\u2212= r jQeiQQe ))sin(())2cos((( 123 \u03b8\u03b8 ++\u2212 (16) where s is the distance between connecting rod\u2019s C.G and piston, and e= s/l . For vertical acceleration of C.G: QQQacx )2cos(4)sin()cos(( 2 21 2 1 \u03c9\u03b8\u03b1\u03b8\u03c9\u03b8 \u2212\u2212\u2212= QQnQ ))2cos())(cos(()2sin(2 542 \u03b8\u03b8\u03b1\u03b1\u03b8 +\u2212\u2212\u2212\u2212 nQQn r sin()cos( 2 1 ))2cos()(sin( 23 54 2 \u03b8\u03c9\u03b8\u03b8\u03c9 \u2212\u2212 COSQQneQ ))2()((cos)sin())2sin( 2 54 222 1 \u03b8\u03b8\u03c9\u03b8\u03b8 \u2212+ iQQe r )))2cos(( 23 \u03b8+ (17) 88 J"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001297_ems.2008.85-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001297_ems.2008.85-Figure1-1.png",
+        "caption": "Figure 1. Induction machine coupled with speed reducer and mechanical load.",
+        "texts": [],
+        "surrounding_texts": [
+            "Taking into account the modelling of the gear, we present here the simulation results of the induction drive operating under healthy conditions. Indeed, a spectral analysis carried out on the stator current shows, in addition to the harmonics related to the machine functioning, the existence of the frequencies which characterize the gear: the meshing frequency fmesh and its harmonics (Figure 8). If the pinion has 11 teeth and the machine is supplied with a frequency fe, the machine is having 2 pole pairs. These frequencies result from the modulation of the meshing frequency by the supply frequency [7]. (1): Pinion velocity (rad/s), (2): well velocity (rad/s), (3): electromagnetic torque and load torque (Nm). In normal conditions, the three motor supply currents have the same amplitude. A spectral analysis to the current signal reveals the presence of a set of the fundamental frequency related with the existence of stator [15]. The spectral contents of the stator currents in healthy operating mode will be used as reference for the spectral study corresponding to the faulty system in order to determine the default characteristics. III. EFFECTIVENESS OF SPECTRAL ANALYSIS (SA) FOR BEARING DAMAGE DETECTION A simulation study will be conducted to find out whether the current and vibration spectra can be successfully used to identify three types of defects, namely, a localized defect on the inner race, outer race or one of the rolling elements. The selected failure modes are the most common bearing failure cases in practice. It will be shown that Sa is as quite effective as classical method like kurtosis in identifying the localized defects in a rolling element bearing [13]."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002371_s00170-010-2582-x-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002371_s00170-010-2582-x-Figure6-1.png",
+        "caption": "Fig. 6 Coordinate systems in six-axis machine center",
+        "texts": [
+            " [19] adopted Denavit\u2013Hartenberg method to construct a kinematics diagram for a head-tilting type. Lee and She[20] derived an analytical equation for three types of five-axis machines; Jung et al. [21] presented a postprocessor for a typical table-rotating/tilting type. Recently, She, Chang, and Huang [22, 23] established transformations to complete postprocessors for configurations with two non-orthogonal rotary axes. In this paper, a kinematics relation will be derived for a sixaxis machining center with three non-orthogonal rotary axes\u2014 two rotary tables and a swivel head, as Fig. 6 shows. Two independent chains are assumed to indicate machine movement. One is relevant to workpiece controlled by bed (Y-axis), primary rotary table (C-axis), and secondary removable rotary table (A-axis). Another is to the cutter controlled by column (Z-axis), transverse head (X-axis), and swivel spindle (B-axis). Before deriving KTA, a specific identification is defined to show the correlation between two coordinate systems. AsFig. 5 Coordinate systems in swivel spindle and tool shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003972_icra.2011.5980334-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003972_icra.2011.5980334-Figure3-1.png",
+        "caption": "Fig. 3. Independent contact regions on a parallelepiped with Qr = 0.05: a) Minimal ICRS, min = 0.1; b) Nominal ICRS, nom = 0.2. Note that the higher the friction coefficient, the larger the ICRi obtained.",
+        "texts": [
+            " Note that diminishing may potentially lead to a situation where the force closure property for the starting grasp cannot be guaranteed any longer. If this is the case, then the computation of ICRS using Algorithm 1 will lead to an empty set of ICRS. The minimal ICRS (if they exist) allow an FC grasp despite any variation of , i.e. they are the most secure ICRS to grasp the object. If at least one finger is outside its ICRmin, then getting an FC grasp cannot be guaranteed due to friction uncertainty. As an example, Fig. 3 shows the computation of the ICRSnom and ICRSmin for a parallelepiped, with = 0.2, = 2 and frictional cones linearized with m = 8 sides. The real ICRS must lie in the ambiguity zone, i.e. somewhere between the ICRS nominal and minimal. The procedure for ICRS computation presented in Section II-C can also be extended to consider soft finger contacts by changing the set W of primitive contact wrenches. The soft finger contact model assumes a finite contact area between the object and the finger, therefore allowing the application of a moment around the local surface normal"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001022_j.jsv.2008.05.023-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001022_j.jsv.2008.05.023-Figure2-1.png",
+        "caption": "Fig. 2. The contact classification of gears: (a) a pair of gears, (b) forward impact and (c) backward impact.",
+        "texts": [
+            " The relative displacement, S, between gears can be expressed as follows: S \u00bc \u00f0Sp \u00fe Ep\u00f0yp\u00de\u00de \u00f0Sg \u00fe Eg\u00f0yg\u00de\u00de (4) where Ep and Eg are the total irregularities of the pinion and gear teeth, respectively, and can be expressed as the error of the tooth pitch on the circumference: Ep\u00f0yp\u00de \u00bc epyp=2p (5) Eg\u00f0yg\u00de \u00bc egyg=2p (6) The pitch error ep, eg (ep, egp90 mm) between the teeth is based on statistical data from previous measurement experiments. Backlash B and displacement S are used to describe the contact-impact behavior between the gears. In this study, the contact-impact behavior is considered as follows (Fig. 2): (i) SXB, the pinion impacts the gear on the forward side of the pinion tooth, and this is called forward impact (Fig. 2b). ARTICLE IN PRESS Q. Gao et al. / Journal of Sound and Vibration 319 (2009) 463\u2013475466 (ii) Sp B, the gear impacts the pinion on reverse side of the pinion tooth, and this is called reverse impact (Fig. 2c). (iii) B4S4 B, the pinion does not contact the gear. Since shafts vibrate in the axial and lateral direction, it is considered that backlash B changes dynamically, and it can be written as B \u00bc B0 \u00fe DB (7) ARTICLE IN PRESS Q. Gao et al. / Journal of Sound and Vibration 319 (2009) 463\u2013475 467 where B0 is the initial backlash in the respective direction on the circumference of the gears. DB is the change value of backlash: DB \u00bc \u00f0Up Ug\u00de nB (8) where Up and Ug are the motion displacement vectors of the pinion and the gear"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000207_s00542-006-0309-6-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000207_s00542-006-0309-6-Figure1-1.png",
+        "caption": "Fig. 1 Two different 1-in. disk drive enclosures",
+        "texts": [
+            "edu form factor hard disk drives using two different enclosures, the so-called thin enclosure and the thick enclosure. First, we perform modal analysis for the two enclosures and the corresponding disk drive models. Then, we investigate the response of the head disk interface to applied in-plane and out-of-plane vibrations for both models. In addition, we also study the effect of head slap due to non-operational shock. Finally, we investigate the effect of texture on the air bearing surface of the slider to operational shock. 2 Numerical results 2.1 Modal analysis Figure 1 shows the top and front views of the two designs of 1-in. disk drive enclosures investigated in this study. The enclosure on the left is described in this paper as the thick enclosure, while the enclosure on the right is described as the thin enclosure. The thick enclosure has a recessed area to hold the disk/spindle assembly and the actuator base. Modal analysis of the enclosures was performed first for the enclosures alone and then for the assembled disk drives including the head suspension assembly, the spindle motor, and the disk",
+            " The stiffness value used for these springs was 40 N/mm. Similarly, two springs with stiffness values of 10 N/mm were positioned in the middle of the inner and the outer rail of the slider to approximate the roll stiffness. To simulate the vibration response for the two 1-in. hard disk drive models, a random displacement between 1 and 10 kHz frequency range was created and applied to the hard disk enclosure (Fig. 3). For the inplane vibration analysis, the displacement was applied along the edge of the enclosure [x or y directions, (Fig. 1)]. For the out-of-plane vibration analysis, the displacement was applied at the base of the enclosure [z direction; Fig. 1]. Figures 4 and 5 show the relative on-track and offtrack displacement at the trailing edge center of the slider for externally applied in-plane vibrations. Ontrack vibrations are the relative displacements of the slider trailing edge center in the circumferential direction of the disk. Off-track vibrations are the relative displacements of the slider trailing edge center in the radial direction. Figure 4 shows the displacement time history for the 1-in. HDD model with the thin enclosure and Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.1-1.png",
+        "caption": "Fig. 1.1. Characteristic shape factors (indicated by points and shaded areas) of tyre or axle characteristics that may influence vehicle handling and stability properties. Slip angle and force and moment positive directions, cf. App.1.",
+        "texts": [
+            "hapter 1 TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY This chapter is meant to serve as an introduction to vehicle dynamics with emphasis on the influence of tyre properties. Steady-state cornering behaviour of simple automobile models and the transient motion after small and large steering inputs and other disturbances will be discussed. The effects of various shape factors of tyre characteristics (cf. Fig. 1.1) on vehicle handling properties will be analysed. The slope of the side force Fy vs slip angle a near the origin (the cornering or side slip stiffness) is the determining parameter for the basic linear handling and stability behaviour of automobiles. The possible offset of the tyre characteristics with respect to their origins may be responsible for the occurrence of the so-called tyre-pull phenomenon. The further non-linear shape of the side (or cornering) force characteristic governs the handling and stability properties of the vehicle at higher lateral accelerations",
+            " When the wheel motion deviates from this by definition zero slip condition, wheel slip occurs that is accompanied by a build-up of additional tyre deformation and possibly partial sliding in the contact patch. As a result, (additional) horizontal forces and the aligning torque are generated. The mechanism responsible for this is treated in detail in the subsequent chapters. For now, we will suffice with some important experimental observations and define the various slip quantities that serve as inputs into the tyre system and the moment and forces that are the output quantities (positive directions according to Fig. 1.1). Several alternative definitions are in use as well. In Appendix 1 various sign conventions of slip, camber and output forces and moments together with relevant characteristics have been presented. For the freely rolling wheel the forward speed Vx (longitudinal component of the total velocity vector V of the wheel centre) and the angular speeA of revolution s can be taken from measurements. By dividing these two quantities the so-called effective rolling radius re is obtained: V x r - (1.1) e $c~ o Although the effective radius may be defined also for a braked or driven wheel, we restrict the definition to the case of free rolling",
+            " During braking, the fore and aft slip becomes negative. At wheel lock, obviously, x = - 1. At driving on slippery roads, x may attain very large values. To limit the slip to a maximum equal to one, in some texts the longitudinal slip is defined differently in the driving range of slip: in the denominator of (1.2) s is replaced by s This will not be done in the present text. Lateral wheel slip is defined as the ratio of the lateral and the forward velocity of the wheel. This corresponds to minus the tangent of the slip angle a (Fig. 1.1). Again, the sign of a has been chosen such that the side force becomes positive at positive slip angle. V tan a - - __LY ( 1 . 3 ) V X 4 TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY The third and last slip quantity is the so-called spin which is due to rotation of the wheel about an axis normal to the road. Both the yaw rate resulting in path curvature when a remains zero, and the wheel camber or inclination angle 7 of the wheel plane about the x axis contribute to the spin. The camber angle is defined positive when looking from behind the wheel is tilted to the right",
+            " The forces Fx and Fy and the aligning torque M z are results of the input slip. They are functions of the slip components and the wheel load. For steadystate rectilinear motions we have in general: F - F ( x , a , 7 , F ) , F y - F ( x , a , 7 , F ) , M z - M z ( X , a , 7 , F z ) (1.4) The vertical load F z may be considered as a given quantity that results from the normal deflection of the tyre. The functions can be obtained from measurements for a given speed of travel and road and environmental conditions. Figure 1.1 shows the adopted system of axes (x, y, z) with associated positive directions of velocities and forces and moments. The exception is the vertical force F z acting from road to tyre. For practical reasons, this force is defined to be positive in the upward direction and thus equal to the normal load of the tyre. Also D (not provided with a y subscript) is defined positive with respect to the negative y axis. Note, that the axes system is in accordance with SAE standards (SAE J670e 1976). The sign of the slip angle, however, is chosen opposite with respect to the S AE definition, cf",
+            " The diagrams include the situation when the brake slip ratio has finally attained the value 100% (x = - 1) which corresponds to wheel lock. The slopes of the pure slip curves at vanishing slip are defined as the longitudinal and lateral slip stiffnesses respectively. The longitudinal slip stiffness is designated as CF,,. The lateral slip or cornering stiffness of the tyre, denoted with C F,,, is one of the most important property parameters of the tyre and is crucial for the vehicle's handling and stability performance. The slope of minus the aligning torque versus slip angle curve (Fig. 1.1) at zero slip angle is termed as the aligning stiffness and is denoted with CM~. The ratio of minus the aligning torque and the side force is the pneumatic trail t (if we neglect the socalled residual torque to be dealt with in Chapter 4). This length is the distance behind the contact centre (projection of wheel centre onto the ground in wheel plane direction) to the point where the resulting lateral force acts. The linearised force and moment characteristics (valid at small levels of slip) can be represented by the following expressions in which the effect of camber has been included: F - CFX M z - -CM~a +CM~Y (1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000521_s11012-006-9049-z-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000521_s11012-006-9049-z-Figure1-1.png",
+        "caption": "Fig. 1 (a) The normal nutation of a spinning wheel starting from A when stationary, then falling to acquire maximum horizontal velocity at B and then rising to C where it is again stationary, momentarily. (b) If the pivot at D is lifted while the wheel is moving from B to C, the nutation ceases. Normal precession commences at C without further nutation, so some horizontal KE has converted to potential energy thereby assisting the lifting force",
+        "texts": [
+            " First of all, let there be a spinning wheel of mass M mounted on a horizontal axle, precessing at a velocity vp in the normal way under gravitational torque. Now, apply a horizontal impulse to the wheel which quickly increases its horizontal velocity to 2vp. This is the maximum nutation velocity and causes the gyroscopic nutation phenomenon, such that the wheel will rise to convert its kinetic energy [1/2M(2vp)2] into gravitational potential energy, then fall to recover this KE in the well-understood manner, see Fig. 1a, French [11], and Stephenson [12]. However, if the axle support pivot is lifted carefully to keep the axle horizontal, from the position (B) where the wheel has maximum horizontal velocity, the nutation process will cease and become normal precession at velocity vp again at C, see Fig. 1b. Thus the original smooth precession condition may be re-established without any loss of the impulse energy; but the spinning wheel is now higher, at increased gravitational potential energy. Consequently, the increase in potential energy (Mgh), and the average vertical lifting force FV over height h, must be related to the impulse energy by an energy equation: Mgh = FVh + { (1/2) M ( 2vp )2 \u2212 (1/2) M ( vp )2 } . (1) That is, the overall increase in potential energy is equal to vertical lifting work done at the pivot, plus the horizontal impulse energy input at the wheel",
+            " If the horizontally pulling force is reversed, then the gyroscope falls rapidly, thereby exerting zero weight on the load-cell. This is just what one would expect when a precessing gyroscope is temporarily prevented from precessing by an obstacle. Ideally, a programmable robot with servocontrol to keep the gyroscope in horizontal alignment, would allow a thorough study of the weight-reduction phenomenon. In particular, it would be advantageous if the position of maximum KE to PE conversion rate could be maintained between B and C of Fig. 1b, by introducing pulses of KE more frequently without waiting for the full relaxation period (t2\u2013t1). Furthermore, returning the axle pivot to the centre of motion might allow a greater range of useful forced precession. Then two gyroscopes operating coherently either side of the centre would be interesting. Weight-reduction of a spinning wheel has been measured using a load-cell, during a short period of enforced precession and simultaneous lifting. It is evident that applied horizontal work input is converted into gravitational potential energy, resulting in weight-reduction of 8% typically"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure3.25-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure3.25-1.png",
+        "caption": "Figure 3.25 Fuel temperature control system schematic",
+        "texts": [
+            " In the example presented the short cut method gave conservative solutions with a gain margin of almost 20 dB and phase margin of over 50 degrees. While this method demonstrates a fairly conservative rule of thumb, designing for specific margins based on pre-determined system performance drivers will require a complete open loop response analysis. The short cut method, however, is an excellent place to start. To demonstrate the application of lead\u2013lag compensation we will use a fuel temperature control system example. This system, shown in the schematic diagram of Figure 3.25 serves to control the temperature of the fuel that is fed to a gas turbine engine by mixing cold fuel from the storage tanks with hot fuel from heat exchanger discharge. Heat exchangers using fuel as a cooling medium are often used as a heat sink for aircraft systems that generate substantial amounts of heat such as hydraulics and avionics. By maintaining the engine feed fuel at a maximum allowable value, the availability of cold fuel as a heat sink is maximized. In the example schematic an electronic controller compares the temperature set point with the measured engine fuel feed temperature"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003873_s00707-012-0647-7-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003873_s00707-012-0647-7-Figure2-1.png",
+        "caption": "Fig. 2 Mapping domains",
+        "texts": [
+            " Utilizing the methods of complex variables [28\u201331] and applying the theory of conformal mapping involving elliptic functions, these functions could be obtained. 2.3 The mapping function of the ellipse into the unit circle Applying the Riemann mapping theorem [32], which states that \u201cAny simply-connected domain D can be mapped conformably onto the interior of the unit circle,\u201d the function \u03b6 = w(z) that maps the interior Dz of an ellipse onto the interior D\u03b6 of the unit circle can be deduced by means of the elliptic functions, see Fig. 2. Then the interior of the ellipse x2 a2 + y2 b2 = 1, a2 \u2212 b2 = 1, z = x + iy, (11) is mapped onto the unit circle |\u03b6 | < 1 by the function \u03b6 = \u03c9(z) = \u221a k(\u03b3 )sn ( 2k \u03c0 sin\u22121 z; \u03b3 ) (12) where \u03b3 = ( a \u2212 b a + b )2 (13) and k2(\u03b3 ) = 16\u03b3 \u221e\u220f n=1 [ 1 + \u03b3 2n 1 + \u03b3 2n\u22121 ]8 . (14) In this mapping, the foci of the ellipse correspond to the points \u03b6 = \u00b1\u221a k(\u03b3 ). For a full proof of function (12), reference may be made to [26]. For a general ellipse a2 \u2212 b2 = f 2, then (a/ f )2 \u2212 (b/ f )2 = 1, where f is the focal distance"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003927_s10846-013-9930-7-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003927_s10846-013-9930-7-Figure1-1.png",
+        "caption": "Fig. 1 Inertial reference frame, FE, body fixed reference frame, FB, and the wind triangle",
+        "texts": [
+            " Wind is of great interest when deriving the equations of motion of an airplane since it directly affects the relative velocity between the vehicle and the air and therefore the amount of the produced lift and drag. Both longitudinal and lateral dynamics are sensitive to wind perturbations, nevertheless, only the lateral-directional control is of primary interest in this study. Hence, only the lateral equations of motion in a non-steady atmosphere will be briefly introduced in the following. The rate of change of the translational position is expressed in terms of an inertial frame, FE, whose axes are according to Fig. 1, and it is given by r\u0307g = Vg where rg is the position of the object relative to FE and Vg is its inertial velocity, usually called ground speed. The speed relative to the ground can be further expressed in terms of air-relative and wind velocities from the wind triangle equation where Va is the airplane velocity with respect to the air, usually called airspeed, and Vw is the velocity of the wind in the inertial reference. There is not direct measurement of the airspeed from a ground position and this is computed by using sensors placed on board. Consequently, its components are expressed in the directions of the body-fixed reference frame, FB, whose origin coincides with the vehicle\u2019s center of mass and whose axes are according to Fig. 1. Let us consider that the airplane is stabilized to fly in level flight. Thus, it can be concluded that the airplane velocity along with the angle of attack, sideslip, pitch and roll angles vary slowly compared to the other parameters, and their time derivatives can be neglected in the flight dynamics. Let BB define the complete transformation from FE to FB BB = ( cos \u03c8 sin \u03c8 \u2212 sin \u03c8 cos \u03c8 ) Thus, it yields r\u0307g = BT BVa + Vw Let xN and yE be the components of the inertial position in the x-axis (North) and in the y-axis (East), respectively, u and v the components of the air relative velocity in FB and wN and wE the headwind and the crosswind relative to FE"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.6-1.png",
+        "caption": "Fig. 3.6. Anthropomorphic arm at a shoulder singularity",
+        "texts": [
+            "2, the determinant does not depend on the first joint variable. For a2, a3 = 0, the determinant vanishes if s3 = 0 and/or (a2c2 + a3c23) = 0. The first situation occurs whenever \u03d13 = 0 \u03d13 = \u03c0 meaning that the elbow is outstretched (Fig. 3.5) or retracted, and is termed elbow singularity . Notice that this type of singularity is conceptually equivalent to the singularity found for the two-link planar arm. By recalling the direct kinematics equation in (2.66), it can be observed that the second situation occurs when the wrist point lies on axis z0 (Fig. 3.6); it is thus characterized by px = py = 0 and is termed shoulder singularity . Notice that the whole axis z0 describes a continuum of singular configurations; a rotation of \u03d11 does not cause any translation of the wrist position (the first column of JP is always null at a shoulder singularity), and then the inverse kinematics equation admits infinite solutions; moreover, motions starting from the singular configuration that take the wrist along the z1 direction are not allowed (see point b) above)",
+            "77) is not cancelled any more and nothing can be said about its sign. This implies that asymptotic stability along the trajectory cannot be achieved. The tracking error e(t) is, anyhow, norm-bounded; the larger the norm of K, the smaller the norm of e.14 In practice, since the inversion scheme is to be implemented in discrete-time, there is an upper bound on the norm of K with reference to the adopted sampling time. Example 3.4 Consider the anthropomorphic arm; a shoulder singularity occurs whenever a2c2 + a3c23 = 0 (Fig. 3.6). In this configuration, the transpose of the Jacobian in (3.38) is JT P = [ 0 0 0 \u2212c1(a2s2 + a3s23) \u2212s1(a2s2 + a3s23) 0 \u2212a3c1s23 \u2212a3s1s23 a3c23 ] . P , if \u03bdx, \u03bdy and \u03bdz denote the components of vector \u03bd along the axes of the base frame, one has the result \u03bdy \u03bdx = \u2212 1 tan \u03d11 \u03bdz = 0, should be increased to reduce the norm of e as much as possible. implying that the direction of N (JT P ) coincides with the direction orthogonal to the plane of the structure (Fig. 3.13). The Jacobian transpose algorithm gets stuck if, with K diagonal and having all equal elements, the desired position is along the line normal to the plane of the structure at the intersection with the wrist point"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.26-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.26-1.png",
+        "caption": "Fig. 1.26. The automobile subjected to longitudinal forces and the resulting load transfer.",
+        "texts": [
+            " When the slip angles become larger, the forward speed u may no longer be considered as a constant quantity. Then, the system is described by a third-order set of equations. In the paper (Pacejka 1986) the solutions for the simple automobile model have been presented also for yaw angles > 90 ~ When the vehicle is subjected to longitudinal forces that may result from braking or driving actions possibly to compensate for longitudinal wind drag forces or down or upward slopes, fore and aft load transfer will arise (Fig. 1.26). The resulting change in tyre normal loads causes the cornering stiffnesses and the peak side forces of the front and rear axles to change. Since, as we assume here, the fore and aft position of the centre of gravity is not affected (no relative car body motion), we may expect a change in handling behaviour indicated by a rise or drop of the understeer gradient. In addition, the longitudinal driving or braking forces give rise to a state of combined slip, thereby affecting the side force in a way as shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002532_detc2012-70485-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002532_detc2012-70485-Figure1-1.png",
+        "caption": "FIGURE 1. Omnicopter MAV Schematic",
+        "texts": [
+            " Their small size provide for low acoustic signatures and radar cross-sections that are ideal \u2217Address all correspondence to this author. for stealth operations [1]. In the last few years, MAVs in the tri-coptor [2\u20134] and quadrotor [5\u201310] configurations have been highlighted in many papers. There has also been some recent work on unique MAV configurations that use gyroscopic moments for attitude control [1, 11, 12]. In the first part of this paper, we will present our novel VTOL (Vertical Take-Off and Landing) MAV design called the Omnicopter (Fig.1) along with its dynamic model based on the Euler-Lagrange formalism. The system model is represented in state space using quaternion and angular velocity in the body frame as state variables. Different control techniques have been applied to the tricopter and quadrotor MAV configurations, such as PID, LQR, 1 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/75895/ on 07/30/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use H\u221e and sliding mode control [5, 7\u20139]",
+            " In the second part of this paper, we focus on the attitude control of the Omnicopter for a special operating case that is similar to that of the quadrotor, also using feedback linearization. Based on the Omnicopter nonlinear model, we design a quaternion-based attitude control algorithm. The zero dynamics problem is also analyzed. Finally, simulations are carried out and results presented to prove the feasibility of the proposed controller. The major conclusions of the paper are then drawn and directions for future work are presented. A schematic of the Omnicopter configuration MAV is shown in Fig. 1. Drawing inspiration from omnidirectional wheels, the Omnicopter design allows for agile movements in any planar direction with fixed (zero) yaw, pitch and roll angles. It has five propellers: two fixed major coaxial counter-rotating propellers in the center used to provide most of the thrust and adjust the yaw angle, and three adjustable angle small ducted fans located in three places surrounding the airframe to control its roll and pitch. Unlike quadrotor or typical trirotor MAVs [2, 3], the Omnicopter has different motion principles and control modes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002660_phm.2012.6228904-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002660_phm.2012.6228904-Figure2-1.png",
+        "caption": "Fig. 2. Different crack levels in the gears",
+        "texts": [
+            " Note that different than MLM, which estimate a set of iJ for every category except the baseline, there is just one equation for eLM: if Xik increases by one, then all transformed cumulative probabilities increase by iJk' Thus, this model is more restricted than a multinomial logit or a hierarchical log it model; by focusing on the cumulative probabilities, we can postulate a single effect of the predictor. For more details, please refer to [6]. III. RE SULT S The dataset we use here is from the experimental system of Lei and Zuo [3], which contains 3 levels of gear conditions (no crack, referred to as 0, and two developing cracks, 25% and 50%, as shown in Figure 2), 3 different loads (no load, half the maximum load and maximum load), 4 different motor speeds (1200rpm, 1400rpm, 1600rpm and 1800rpm) under identical operating conditions, with three data samples collected. (The maximum load was calculated when the maximum stress was less than the allowable stress.) The sampling frequency is 5210 Hz and there are 8192 sampling points. Therefore, we obtain 36 data samples at each crack level and 108 samples altogether. First we compare our feature selection method with that of Lei and Zuo [3]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001057_20080706-5-kr-1001.01564-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001057_20080706-5-kr-1001.01564-Figure2-1.png",
+        "caption": "Fig. 2. Experimental setup.",
+        "texts": [
+            " (25) Therefore, following (23) we can show that there exist positive c0, c1 and c2 such that ||e(t)||2 \u2264 c0e \u2212\u03b7t||e(0)||2 + c1w\u0304 2 + c2\u2206 2, \u2200t \u2265 0. (26) where \u03b7 = \u03bbmin(I \u2212 \u03c4P )/\u03bbmax(P ). We can thus take \u00b5 = \u221a c0\u03b12 + c1w\u03042 + c2\u22062 (27) for (22). This section presents the application of the reset estimator to a linear motor positioning system with optical encoder. Simulation and experimental results are shown to verify the effectiveness of the reset state estimator (RSE) over the standard state estimator (SSE) without reset actions. Fig. 2 shows the experimental setup. The linear motor driven stage has a 0.5 m travel range, a mounted optical encoder of 1 \u00b5m resolution , and a power amplifier. The linear motor is modeled from its physical parameters and the model in state space is given by [ y\u0307 v\u0307 ] = [ 0 1 0 \u2212a ] [ y v ] + [ 0 b ] u, yq = Q(y), (28) with a = 7.5398, b = 1.5 \u00d7 107, and y is the position (in \u00b5m), v is the velocity of the stage (in \u00b5m/s), respectively. The block diagram of the linear motor control system is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure13-1.png",
+        "caption": "FIG. 13. Scheme of flat air-bearing test rig.",
+        "texts": [
+            " In the event of air cut-off there was an allowable run-down time of over 30 sec, which was more than adequate to allow the machine to stop all movements. Before manufacturing this spindle it was decided to gain some experience on a rig, to determine convenient sizes for jets and to optimize stiffness. The experience was useful since it was possible to assess the effect on an air bearing of various manufacturing errors, and to know that the method of manufacturing jets was satisfactory for the limits chosen. A flat rig (Fig. 13) was designed to simulate the conditions of air flow in a lightly loaded journal bearing with a single row of jets in each bearing. The shim flapper seals ensured the condition of zero circumferential flow but were adjusted to have negligible mechanical restraint on the air gap. Grinding Machine Spindles 467 The bearing faces of the rig were scraped to obtain parallel surfaces but it was still found that the results for a bearing clearance less than 0-0002 in. were not very reliable. However for the sizes of jets used this was in the region of air film collapse so that this point was not serious"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003909_j.elecom.2013.10.029-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003909_j.elecom.2013.10.029-Figure1-1.png",
+        "caption": "Fig. 1. (A) Sketch of the plane-band-recessedmicrodisk array electrodes assemblymicrofabricated and used in thiswork; (B) laser confocal image of the device; (C) sketch of the diffusionelectrochemical paths followedby the various specieswhen the assembly is operated as described in the text in thepresence of a sample containing dopamine (DA) and ascorbic acid (AA); DAox represents the quinonic form of oxidized dopamine and AAox the non-electroactive hydrated oxidation product of AA.",
+        "texts": [
+            " Based on the fact that our electrode configuration is composed of three working electrodes, lithography technique is selected to fabricate the electrode arrays. Compared with plane-recessed microdisk array electrodes [3,21], the only change consisted in the additional implantation through the same technique of a conductingAu layer located in between the recessed disks and the top plane and separated from these two elements by a polyimide insulating layer. This intermediate layer was perforated so that it appeared as an array of ring band electrodes located in the wells (Fig. 1A). For achieving the perforation, a layer of polyimide was coated, which was followed by photolithography, lift-off of Cr/Au layer, and etching polyimide by inductively coupled plasma. The vertical height distance between top plane and recessed disk electrodes was measured by profilometer (Dektak3 Series). All electrochemical measurements were carried out on a CHI814A potentiostat (Chenhua Corp., China). A standard polypropylene pipette tip was used as the electrochemical cell. A saturated calomel electrode (SCE) and a gold wire were used as reference electrode and counter electrode, respectively. Due to the limitation of the volume of pipette tip, the SCE was placed in a separated cell and connected with the pipette tip via a KNO3 agar salt bridge. The plane-band-recessed microdisk array electrodes were characterized by laser confocal microscopy and profilometer. Fig. 1A shows a sketch of the array, which is composed of a plane electrode and an array of microcavities. The Au microdisk electrodes were located at the bottom of the cylindrical microcavities performed through the Au top plane and its underlying insulator, and the ring-band electrodes (0.2 \u03bcm width) located between the bottom and top Au electrodes. A laser confocal image of the device is shown in Fig. 1B, which demonstrates that the array componentswere perfectly aligned as designed. The disk electrodes had a ca. 13 \u03bcm diameter and their centers were separated by ca. 72 \u03bcm. In addition, the vertical distances within the microcavities were measured with a profilometer. The depth of microcavities, the distance between disks and ring-band electrodes, top-plane and ring-band electrodes in each well were respectively 8.6, 4.1 and 4.1 \u03bcm, to ensure the overlapping of diffusion layers generated by the three electrodes during operation (Fig.1C). The diffusion-electrochemical pattern shown in Fig. 1C is generated as follows. First, during a rest period a sufficiently anodic potential is applied to all electrodes to oxidize both DA and AA, which creates a steady-state AA and DA-depleted microenvironment in the well and above through converting DA into its quinone form and AA into its non-electroactive hydrated 2e-oxidation product. Then, during the measuring phase, the recessed microdisk electrodes array potential was swept negatively while the plane and band electrodes remained poised at the rest potential"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000918_s11465-008-0013-6-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000918_s11465-008-0013-6-Figure1-1.png",
+        "caption": "Fig. 1 Demonstration of reconfigurable 6-DOF modular robots",
+        "texts": [
+            " Thus, the inverse kinematics of reconfigurable modular robots are converted to transform the POE formula to standard subproblems by using the appropriate simplification method and transformation. Translated from Chinese Journal of Mechanical Engineering, 2006, 42(8): 210\u2013214 [\u8bd1\u81ea: \u673a\u68b0\u5de5\u7a0b\u5b66\u62a5] Jie ZHAO, Weizhong WANG (*), Yongsheng GAO, Hegao CAI Robotics Institute, Harbin Institute of Technology, Harbin 150001, China E-mail: wangwzh@hit.edu.cn 2 POE formula and its reduction techniques 2.1 Product-of-exponentials formula The geometries and DOFs of reconfigurable robots vary; some typical 6-DOF configurations are shown in Fig. 1 [10] where revolute joints and prismatic joints are represented by cylinders and rectangular, respectively. The product-of-exponentials formulation of robot kine- matics of an n-DOF serial-type robot has the form [10]: gst( h )~ exp j\u03021h1 exp j\u03022h2 . . . exp j\u0302nhn gst( 0 ), \u00f01\u00de where j\u0302i ( i~1, 2, . . . , n): are the joint twists of a instan- taneous motion, exp j\u0302ihi represents a rigid motion. gst(0) and gst(h) are the initial and final poses (positions and orientations) of the end effector. j : ~(j\u0302)_[R6 is defined as the twist coordinates for twist j\u0302"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001163_20080706-5-kr-1001.02201-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001163_20080706-5-kr-1001.02201-Figure1-1.png",
+        "caption": "Fig. 1. Engagement Geometry",
+        "texts": [
+            " The simulation results show that the MPSAG technique is quite promising to address the problem preserving the nonlinear kinematics of engagement. It is observed that the new nonlinear guidance law proposed in the paper leads to the following advantages as compared to the existing linearized optimal guidance law :(i)Latax calculation is carried out using nonlinear dynamics and (ii) singularity at the end of guidance command for first order delay system is completely eliminated. 978-3-902661-00-5/08/$20.00 \u00a9 2008 IFAC 13016 10.3182/20080706-5-KR-1001.1450 Fig. 1 shows the engagement geometry for a stationary and slowly moving target. The missile is moving with a constant speed and the target under consideration is assumed to be stationary. The guidance command is the lateral acceleration (latax), which is the acceleration normal to the velocity vector (Ryoo et al. (2005)). The equation of motion for this homing missile problem is formulated both for lag free and first order autopilot delay system Impact angle, for a stationary target is the missile flight angle at the time of interception (Ryoo et al"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002282_978-90-481-9262-5_35-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002282_978-90-481-9262-5_35-Figure5-1.png",
+        "caption": "Fig. 5 A trajectory in which point P crosses the segment 8\u2019-8\u201d.",
+        "texts": [
+            " Figure 4f also shows the results of the classification process given in Section 3 applied to one point for each one of the returned boxes. It is worth mentioning that the segment 8\u2032-8\u2032\u2032 was erroneously marked as an interior barrier in [16], while we detect it as a singularity of type non-barrier. This segment corresponds to point P tracing a circle around point B, when l1 is fixed to its lowest value \u221a 2, while keeping the platform aligned with leg 1. The result in [16] must be erroneous, because P can really cross this segment from any of its two sides, as shown in Fig. 5. The platform can start from a position where P is to the right of the segment (Fig. 5a), then slide down along line L until it hits the segment (Fig. 5b), and, locking l1 and l2 to their values in this configuration, finally perform a rotation about point A by actuating l3 (Fig. 5c). This paper has introduced a new approach to compute workspace boundaries of general multi-body systems. A principal advantage of the method is its ability to converge to all boundary points, as discussed in the paper. Previous methods for the same purpose cannot ensure this property, since they are based on continuation, which requires the availability of one point for each connected component of the sought boundary, and no previous work on workspace analysis has shown how to compute all of such points in general, to the authors\u2019 knowledge"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002017_10426914.2010.496126-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002017_10426914.2010.496126-Figure13-1.png",
+        "caption": "Figure 13.\u20143D shape of the turbine component.",
+        "texts": [
+            "56 50 41 48.67 \u22123.45 Height of small circle (h2) 2.81 2 08 2.16 3.85 Diameter of small circle (d2) 13.82 12 11 11.23 \u22127.27 aRelative error = (simulation result \u2212 experimental result)/(experimental result). bName of the dimension; see Fig. 14. Figure 17.\u2014Plastic strain contour in the X-axis of the turbine component after cold isostatic pressing. The CIP process of a turbine component was simulated as an example. Because the turbine component is symmetric about the 1\u20133 plane as shown in the 3D part of Fig. 13, only the upper half of the component was modeled. And because of the axisymmetric structure, the axisymmetric model was used. The cross-section dimensions of the turbine component are shown in Fig. 14. The component after SLS is shown in Fig. 15 and the initial relative density of the part is 0.38. The pressure was 630MPa and was applied to the free surfaces of the turbine component. The meshes before and after CIP are shown in Fig. 16. The part has little deformation in shape but has a large compaction in volume"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002029_j.optlaseng.2010.04.003-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002029_j.optlaseng.2010.04.003-Figure1-1.png",
+        "caption": "Fig. 1. The 3D reconstruction theory of close range photogrammetry.",
+        "texts": [],
+        "surrounding_texts": [
+            "The transmission towers have become higher because of the increase of transportation voltage in these years. It makes an estimate of the loading capacity for transmission towers more important. A lot of transmission towers were damaged by the snow disaster in south China, 2008. Most reported studies on loading deformation of transmission towers employed numerical simulation method, which can predict three-dimensional deformation of the structure at low cost, but the predicted result is often far away from the real experiment, due to complexity of the loading process [3\u20137]. Moon drew a conclusion that the simulation results cannot agree with the experimental data well when the loading conditions are complex [1]. The loading capacity of transmission towers cannot be calculated accurately only by numerical simulation. Measuring the field deformation of real-sized model and the loading capacity of transmission towers under different load conditions are necessary. The traditional displacement sensors and resistance strain gauges can measure only one-dimensional deformation in limited measurement range, and it cannot measure the whole field deformation of transmission tower [1\u20135]. Traditional sensors are all contact-type device, install of them is difficult or impossible in many cases, especially when access to the structure is prohibited. In order to solve these kinds of problems, optical measurement has been introduced by more and more researchers. Fraser (2000) [5] pasted some fire-resistant artificial markers made of special ll rights reserved. . material on the transmission tower to measure its displacement during the temperature falling from 11001 to room temperature by optical measurement to obtain the diagram of displacement in each step. Jiang (2007) [7]proposed a new scheme controlled by meshwork, while measuring the cement beams deformation of large bridge with loads to decrease the markers and workload when measuring. Xu Fang et al. used digital photogrammetry to measure the deflection of steel structure. Huang measured windinduced liberation of a high altitude cantilever transmission tower at the scene by photogrammetry to solve the problem that measuring the inaccessible object dynamically [7\u201312,16]. In this paper, a non-contact 3D optical static deformation measurement system called XJTUSD is developed by a Xi\u2019an Jiaotong University in China, in order to monitor the 3D deformation of realsized transmission tower during loading test. The key technologies of close range industrial photogrammetry are studied, including the disposition of markers, the disposition of camera stations, deformation of single step, measurement coordinates alignment of different steps, the registration method for corresponding points in multistep, calculation and display for the deformation. The deformation of the whole transmission tower with different loads is obtained which is helpful for further quantitative analysis."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000294_tac.1974.1100518-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000294_tac.1974.1100518-Figure6-1.png",
+        "caption": "Fig. 6. Disturbance response constraint for linear and EEAS designs on Nichols ch&.",
+        "texts": [
+            ": THEORY FOR EEaS 105 As noted, in the above high w region IS/ >> 1 is tolerabIe in its effect on T(jw), because the prefilter F in Fig. 1 attenuates the resulting high peaking in L/(1 + L). However, the disturbance response (Fig. 1) C /D = (1 + L)-1 = S, is then &o highly peaked, and there is no equivalent filter available. Even if the parameter ignorance problem dominates, it is necessary to consider the disturbance response, to the extent of adding the constraint In the purely linear design, (10) gives the bound Bh of Fig. 6 (for (K2/K1) = loo), effective ++ w, with eh = 180 - e, a function of y, e.g., 0, = 50\u00b0 if y = 2.3 dB. At small w the parameter ignorance factor dominates, e.g., Be in Fig. 6 which contains no part of B,. There is an intermediate region whose B, contain a portion Bh, e.g., Bd. Eventually Bh applies ++ w 2 oh sufficient,ly large. B,, the equivalent of B, in the EEAS has none of the vertical region UV because of the EEAS zero sensitivity property (10). In the linear case Lopt decreasa along UV in Fig. 6 at an average rate of (e, 12/180) dB per octave, over the interval (wo,w,) in Fig. 2. When L reaches point W its phase lag can be increased, permitting a larger -dlLI/ dw. It has been shown [14] that the optimum L lies precisely on B j at each w1 and follows B, along W X in Fig. 6, but in practice one more often follows approximately the horizontal line W\u2019X\u2019 corresponding to (we,wg) in Fig. 2. In the EEAS the sharp increase in phase lag can occw at a much lower frequency, revealing the great potential saving in loop bandwidth. Application to EEAX Design In order to apply the above technique t.0 the EEAS, it is necessary to mite L, in the form L, = GP, with P, t.he \u201ceffective plant\u201d from the viewpoint of AlnL = AlnP, due to plant uncertainty. Now Lf = 0.5(M/A) KPhGlG2G3(jw) and if Ph has no variation then KIA is constant and A l d \u2019 , = 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001647_9780470630693.ch5-Figure5.21-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001647_9780470630693.ch5-Figure5.21-1.png",
+        "caption": "FIGURE 5.21 The air diffusion CueO-modified electrode [49].",
+        "texts": [
+            "35V, which is close to formal potential of the type 1 site in CueO (0.26V) [50]. Usually the biocathodes rely on dioxygen dissolved in the electrolyte solution; thus themaximumcurrent is limitedbyO2diffusionbecauseof lowsolubility and small diffusion coefficient of oxygen. The O2 diffusion depends on stirring or rotation rate, or flow rate of the electrolyte solution. The air-diffusion-type biocathode was constructed using nanostructured carbon particles, carbon paper, and a binder to glue the particles on paper (Fig. 5.21). 196 BIOCATHODES FOR DIOXYGEN REDUCTION IN BIOFUEL CELLS Following the earliest experiments with nanostructured electrodes modified by laccase [16], different, mainly carbon-based, nanomaterials are employed to improve the heterogeneous electron transfer kinetics on the enzyme-modified electrodes and to increase the catalytic current density. Enzymes adsorb on these materials or are covalently bonded (see section 5.3). Only 10% of the immobilized enzyme molecules take part in the bioelectrocatalytic reaction because of the unfavorable orientation of the adsorbed laccase to the conducting surface [227]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002487_j.precisioneng.2011.02.001-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002487_j.precisioneng.2011.02.001-Figure12-1.png",
+        "caption": "Fig. 12. Portion of a zoom lens design.",
+        "texts": [
+            " T \u03b5 4 4 c t S o m l L c f a L C L a i e T g h d T T he element tilt m, decenter \u03b5m, optical axis tilt o, and decenter o can be calculated with the previously mentioned equations. . Simulation results and discussion .1. Lens preparation In this paper, a design for a 10\u00d7 zoom lens for a digital video amera is chosen for the case study. Most of the optical construcions of high-ratio zoom lenses consist of four or five lens groups. ome groups are fixed while others are movable independently of ne another during zooming. Fig. 12 shows the optical and optoechanical design of the second and third group of a 10\u00d7 zoom ens. There are three lenses in Group II. Lens L5 contacts with lens 6 directly; lens L6 and lens L7 are separated by a spacer S2. All omponents are assembled from the left side of cell C2. There are our lenses in Group III. Lens L8 is a lateral compensating lens nd must be attached from the left side of the cell C3. Lenses 9\u2013L11 and spacer S3 are loaded from the right side of the cell 3. Lenses L9 and L10 are separated by spacer S3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000744_j.mechmachtheory.2008.01.002-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000744_j.mechmachtheory.2008.01.002-Figure1-1.png",
+        "caption": "Fig. 1. Sketch map of dual motor system.",
+        "texts": [
+            " According to the existence conditions [6,7] of the sliding mode, those are [s(k + 1) s(k)]s(k) < 0 and j s\u00f0k \u00fe 1\u00de s\u00f0k\u00de j< n 2 , the control functions can be found to achieve the accurate synchronism and tension control of the dual motor system. Finally, in order to examine the performance of the proposed control scheme, an experimental prototype of the dual motor system is developed in this paper. The experimental results will show the practical capability of the proposed control scheme. To examine the proposed control scheme for the MIMO system, an experimental prototype of dual motor system is developed in this paper. The sketch map of the system is shown in Fig. 1. The leading motor, marked as M1, is assigned as the master motor. In industrial application, such as rolling mill and each machine, the master motor is used to maintain the constant rolling speed in order to make sure the quality of the product. The second motor, marked as M2, is the slave motor usually used as the line putting machine. The control objective is to follow the speed of the master motor and correct the tension of the line. The tension of the line is measured by a linear potential meter"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000827_978-0-585-34652-6_10-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000827_978-0-585-34652-6_10-Figure11-1.png",
+        "caption": "Figure 11. The inverted pendulum system",
+        "texts": [
+            " The chattering in the fuzzy-sliding mode control system is smaller than that in the conventional sliding mode control system 298 In this subsection, we will apply the scheme of FSMC to deal with two dynamic physical systems. One is a highly nonlinear inverted pendulum system, and the other is a two-link robot arm which has not only nonlinearity but also coupling effects. / . Inverted Pendulum System We will illustrate the performance of the FSMC by applying it to an inverted pendulum system as shown in Figure 11. In the inverted pendulum system, a cart which is controlled by an actuactor force/is free to move on a one-dimensional track. A rigid pole is mounted on this cart. The dynamics of the inverted pendulum system can be characterized by two state variables, with x, denoting the angle of the pole with respect to the vertical line, 6, and Xj denoting the angular velocity of the pole. By neglecting the coefficient of friction, the state Eq. can be expressed as follows [12]: gs inxi-cosxi [-;;^xlsmxi-^;;^f^ \"\"' \" f/ 2f-cos2x, where g - the gravity constant, 9"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000614_05698197408981434-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000614_05698197408981434-Figure3-1.png",
+        "caption": "Fig. 3-Sectional asnembly drawing of the test-rig",
+        "texts": [
+            " In order to do this, it is necessary to separate the rib torque from the overall bearing torque. For this purpose, the rib has to be detached from the inner race. The test rig built to do this essentially consists of a tapered roller bearing with outer race (cup) rotating under a thrust load applied by a hydraulic cylinder. The rib has been detached from the inner race and is supported on a circular hydrostatic bearing pad. The test rig has been designed to carry a test thrust load up to 4,000 pounds and is to be operated up to 1,750 rpm. Figure 3 shows the sectional assembly view of the test rig. The detached rib is individually free to rotate and the rib torque is measured off a calibrated strain gage attached onto a lever arm fastened to the alignment bracket. SAE 20 mineral oil is supplied under pressure to the hydrostatic pad to support the thrust load and this same oil is then passed through the test bearing assembly to be used as the lubricant for the raceways and the rib roller end contact. The selection of the test bearings has been based mostly on the adequate overall dimensions of the bearing"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002985_s11768-012-9152-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002985_s11768-012-9152-8-Figure1-1.png",
+        "caption": "Fig. 1 Three-axis frame structure.",
+        "texts": [
+            " Taking advantage of H\u221e con- trol robustness [7], an H\u221e controller is designed for stabilizing loop of tracking and sighting pod platform frames to improve the antijamming capacity and robustness of the system. The application of fuzzy PID controllers in tracking loop is to optimize the dynamic capability of the system, and to improve the tracking precision and the adaptability to parameters changes of the system. Tracking and sighting pod operating at tracking state is a three-degree-of-freedom angular position servo system. It consists of a yaw loop, a pitching loop and a roll loop, as shown in Fig. 1. A photo-electricity detector is connected to the yaw loop. The yaw loop connected to the pitching loop by the yaw-axis can be turned by its axis relative to pitching loop. The pitching loop connected to the roll loop by the pitching-axis can be turned by its axis relative to roll loop. Received 22 July 2009; revised 5 August 2010. This work was supported by the National Defense Key Laboratory, and the Program for Youth Skeleton Teacher in University of Henan Province. c\u00a9 South China University of Technology and Academy of Mathematics and Systems Science, CAS and Springer-Verlag Berlin Heidelberg 2012 The roll loop is the outmost loop; connected to the air-seat by the roll-axis, it can be turned by its axis relative to the air-seat"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001199_0278364907085565-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001199_0278364907085565-Figure4-1.png",
+        "caption": "Fig. 4. Velocity contribution and velocity error vectors computed at the first iteration by (a) VPU 1 and (b) VPU 2.",
+        "texts": [
+            " For such a of problem it is well known that the best solution is represented by a least-squares form q1 1 h# 1 (11) where the pseudo-inverse row vector h# 1 hT 1 hT 1 h1 (12) in (11) plays the same role as the matrix J # in (9). 5. In other words, vectors h1 and h2 constitute the columns of the 2 2 Jacobian matrix of the robot. at CARLETON UNIV on March 16, 2015ijr.sagepub.comDownloaded from Although solution (11) guarantees a minimum-norm error, such an error vector is not generally the null one indeed the resulting end-effector velocity induced by q1 1 that is x1 1 h1 q1 1 (13) is generally (as in the case at hand, see Figure 4(a)) different from the required velocity. Therefore, a velocity error vector (corresponding to the component part of the original task which is still unaccomplished) is originated: 1 1 x1 1 I h1h# 1 (14) and can be easily evaluated by VPU 1 via definition (14). Once computed, vector 1 1 is then sent, as a reference input, to the second VPU, invoking for its corrective contribution (see Figure 4(b)). Then, since VPU 2 acts in the same manner as VPU 1, it also finds out its corresponding scalar joint velocity term by solving a least-squares problem, thus obtaining q2 1 h# 2 1 1 (15) (where pseudo-inverse row vector h# 2 has been obtained by applying (12) to vector h2). Although q2 1 induces a tool velocity term, x2 1, contributing to the execution of the original velocity task, part of , corresponding to the error vector 2 1 1 1 x2 1 I h2h# 2 1 1 (16) still remains unaccomplished. Therefore, with the aim of fulfilling such a residual part, a second iteration, identical to the first that has just concluded, can be started by asking VPU 1 to provide a tool velocity contribution maximally similar to vector 2 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure7-1.png",
+        "caption": "FIGURE 7. The compound pulley (Mechanics 2.12). Drachmann\u2019s drawing from Ms L is on the left (The Mechanical Technology, p. 70); the figure from Heronis Alexandrini opera, vol. II, p. 124, is on the right.",
+        "texts": [
+            " If the weight is divided into two equal parts, each one will balance the other, just like the two weights suspended on the circumference of the larger circles in 2.7. This result is then generalized to cover multiple pulleys and ropes, yielding a precise quantitative relationship: the ratio of the known weight (t iql) to the force (quwwa) that moves it is as the ratio of the taut ropes that carry the weight to the ropes that the moving force (al-quwwat al-muh. arrikat ) moves.35 For example, in Fig. 7, each of the four lengths of rope stretched between the weight Z and the two pulleys on A carries \u00bc the total weight. If we imagine detaching the rightmost section of the weight Z from the sections G\u030cBT. , it will hold those sections in equilibrium. Thus a force equal to \u00bc the total weight of Z, applied at K, balances \u00be the total weight (G\u030cBT. ), and a slightly larger force will move it.36 In the case of the wedge and screw the similarity to the concentric circles is much less clear. Since Heron claims that the screw is simply a twisted wedge (2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000923_bfb0110378-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000923_bfb0110378-Figure4-1.png",
+        "caption": "Fig. 4. The robot 2kTr caxrying its pendulum.",
+        "texts": [
+            " The P V T O L system, the gantry crane and the robot 2k~r (see below) are of this form, as is the simplified planar ducted fan [52]. Variations of this example can be formed by changing the number and type of the inputs [48]. = - u a sin 0 + eu2 cos 0 = Ul cos0 + eu2 s i n 0 - 1 0 z ~ 2 . A fiat output is y = (x - e sin 0, z + e cos 0), see [38] more more details and a discussion in relation with unstable zero dynamics. E x a m p l e 11 ( T h e r o b o t 2kTr o f Eco le des M i n e s ) In [31] a robot arm carrying a pendulum is considered, see figure 4. The control objective is to flip the pendulum from its natural downward rest position to the upward position and maintains it there. The first three degrees of freedom (the angles 01,0~, 03) are actuated by electric motors, while the two degrees of freedom of the pendulum are not actuated. The position P -- ( x , y , z ) of the pendulum oscillation center is a fiat output. Indeed, it is related to the position S = (a,b,c) of the suspension point by (x - a ) ( ~ + g) = ~ ( z - c) (y - b) (2 + g) = i j ( z - c) (x - a) 2 + (y - b) 2 + (~ - ~)'~ = 1 ~ where 1 is the distance between S and P"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000689_09544100jaero206-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000689_09544100jaero206-Figure4-1.png",
+        "caption": "Fig. 4 Co-flowing nozzle (grid arrangement)",
+        "texts": [
+            " A comprehensive experimental and computational fluid dynamics (CFD) analysis had before been performed to prove the feasibility of this concept [35]. Since the classical and most efficient nozzles have been designed with a circular cross section, it was also remarkable to analyse the circular shape for the exhaust nozzle. Here, the engine exhaust was assumed to end to a divergent Coanda surface, which the secondary jet follows in order to vector the primary flow. This model is schematically shown in Fig. 4. A cell centred finite volume technique was used to solve the Reynolds-averaged Navier\u2013Stokes equations (RANS). RANS is potentially a very powerful tool for Coanda wall jet calculations. For these equations, the flow was assumed to be turbulent in all flow-fields and the standard K-\u03b5 model was selected to study the effects of turbulence on flow. It was found that the specified nozzle is capable of vectoring the engine\u2019s exhaust gases up to 29\u25e6. A sample graph of thrust deflection angle for Olympus engine in 78 000 rpm, which is found by CFD analysis and experiments, is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003101_acc.2011.5990996-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003101_acc.2011.5990996-Figure1-1.png",
+        "caption": "Fig. 1. Group of two-link robots",
+        "texts": [
+            " Remark 6: To improve the transient performance, we can incorporate the adaptive redesign and nonlinear damping together [12], as u\u0303i = \u2212\u03c6i\u03b8\u0302i \u2212 \u03b3izi \u2016 \u03c6i \u2016 2 2 , (33) Take derivative of Va in (32) with respect to the trajectories of system (31), and choose ui = u\u0302i + u\u0303i, u\u0302 as the nominal control law (9), and u\u0303i as the redesign (33), we have V\u0307a \u2264 \u2212k3 \u2016 (L \u2297 Im)q \u20162 2 \u2212k4z T z \u2212 k n \u2211 i=1 \u03b3i \u2016 zi \u2016 2\u2016 \u03c6i \u2016 2 +k n \u2211 i=1 zT i \u03c6i\u03b8\u0303i + k n \u2211 i=1 \u03b8\u0303T i \u0393\u22121 i \u02d9\u0303 \u03b8i, for some k3, k4 > 0. Still choose \u02d9\u0303 \u03b8i = \u2212\u0393i\u03c6 T i zi, we have V\u0307a \u2264 \u2212k3 \u2016 (L \u2297 Im)q \u20162 2 \u2212k4z T z \u2212k n \u2211 i=1 \u03b3i \u2016 zi \u2016 2\u2016 \u03c6i \u2016 2\u2264 0. Following the same procedure as the proof of Theorem 2, we can prove the convergence of synchronization. The damping term \u2212\u03b3izi \u2016 \u03c6i \u20162 2 will fortify the adaptive process and smoothen the transient trajectories. We will show this point in the simulation. As shown in Fig. 1, assume the group has four two-link manipulators and the communication topology is a digraph. For convenience, we set the link weights aij = 1, j \u2208 Ni, i, j \u2208 {1, 2, 3, 4}. As shown in Fig. 1(b), the two-link manipulator model [21] has two degrees-of-freedom, the generalized coordinates q = [\u03b81, \u03b82] T . The detailed modeling process is discussed in [16]. We choose the parameters as the Table 5.1 in page 129 of [21] and we set the reference velocity vd = [ 0 2 sin(\u03c0t) ]T , 0 for Angle 1st and 2 sin(\u03c0t) for Angle 2nd. The nominal control law is chosen as (9) with k0 = k1 = 5. The synchronization performance is shown in Fig. 2. In the uncertain model, we perturbed 10% of the model parameters, Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001283_s11029-010-9111-8-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001283_s11029-010-9111-8-Figure4-1.png",
+        "caption": "Fig. 4. Sche matic sketch of a cross-ply rub ber-cord to roid al shell.",
+        "texts": [
+            " We should mark off two im por tant re sults: first, the so lu tion of bend ing prob lems for iso tro pic and com pos ite rings un der the ac tion of a fol lower pres sure is pos si ble with out the use of the in cre men tal ap proach (NStep = 1) and re quires only seven it er a tions, and, sec ond, the num ber of load ing steps does not affect the calculated values of displacements, which confirms the efficiency of the geometrically exact shell elements constructed. Now, let us con sider a four-layer cross-ply rub ber-cord to roid al shell of cir cu lar cross sec tion, loaded with a pres sure p0 uni formly dis trib uted over its in ner sur face (Fig. 4). This shell will be used for mod el ing a di ag o nal tire. The ini tial char ac - ter is tics of the el e men tary rub ber-cord lay ers are [21] EL = 510.45 MPa, ET = 6.91 MPa, GLT = 2.33 MPa, GTT = 1.77 MPa, and nLT = 0.46. Let the shell thick ness be h = 4.8 mm, the thick ness of a rub ber-cord layer hk = 1.2 mm, and the ori en ta tions of rub ber-cord lay ers g gk k= - -( )1 1 , where g = 45\u00b0 and k = 1 4, . As a ref er ence sur face, we as sume the shell midsurface formed by ro ta tion of a part of the cir cum fer ence of ra dius R = 50 mm"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002809_sled.2010.5542802-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002809_sled.2010.5542802-Figure10-1.png",
+        "caption": "Fig. 10. Motors with cost-effective encoder systems (a) LSM motors, equipped with resolvers and capacitive encoders (b) Integrated Drive for lift door application (c) Linear motor with MR sensor to detect the air-gap field of the permanent magnets of a linear motor (prototype)",
+        "texts": [
+            " In the following the authors try to give an outlook, if these methods will be used for a wide variety of applications or will be applied only in special niches. To make this forecast, especially the low-cost encoders have to be considered, which can be an alternative to the sensorless control, namely \u2022 resolvers, \u2022 capacitive or inductive measurements systems, some of them delivering even absolute position information, \u2022 magneto resistive (MR) sensors, \u2022 Hall sensor elements in the motor windings and \u2022 Hall or MR sensors to measure directly the air gap field of the permanent magnets in linear motors. Fig. 10 gives some examples of low-cost encoders applied at LTi. A motor series is offered with resolvers, which are available for prices below 30\u20ac when high numbers are ordered. Resolvers with one pole pair offer even a single turn absolute position with a resolution of 16 bit per turn. These motors can be also equipped with single turn absolute measurement systems, the capacitive encoders SEK of company Sick/Stegmann [12] for approx. 25\u20ac. These encoders offer 16 sine periods per turn, which results in a higher resolution of almost 20 bit per turn. For encoder prices of approx. 50\u20ac these motors can be even delivered with multi turn absolute encoders, so that a reference run at machine start can be avoided. However, for application specific OEM products even much cost-effective solutions are possible. For an integrated door drive (Fig. 10 (b)) position measurement has been developed, which consists of a magneto resistive chip and a permanent magnet, that costs including the electronics on the inverter PCB less than 5\u20ac. Also for linear motors solutions exist with Hall or MR sensors (Fig. 10(c)), that detect directly the air gap flux of the permanent magnets and are in comparison to standard linear encoders less accurate, but really cost effective and robust. As a consequence, the costs cannot be the only argument for applying a sensorless control. On the other hand more and more speed controlled drives are equipped with synchronous instead of asynchronous motors, to reduce energy consumption, size or inertia of the drive system. These drives have been controlled sensorless in the past, either by VF or sensorless vector control, and therefore the customers expect that sensorless operation will be also feasible with the synchronous motors"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure12-1.png",
+        "caption": "FIGURE 12. Combination of compound pulleys to move a weight of 1000 talents with a force of 5 talents. Drachmann\u2019s drawing above (The Mechanical Technology, p. 87) is from Ms B; the figure below is from Heronis Alexandrini opera, vol. II, p. 156.",
+        "texts": [
+            " A similar point follows for the screw, since it is just a twisted wedge; again, it is a crucial assumption that the number of turns of the screw in a given time remains the same (in the text above: \u201cthe turning of the screw many times takes more time than a single turn\u201d). In the case of both wedge and screw, then, it is clear that Heron\u2019s understanding of the phenomenon of slowing up involves a comparison between machines of different mechanical advantage; moreover, this comparison assumes that the moving forces in the two machines travel at the same speed. Such a comparison between moving forces also underlies Heron\u2019s account of slowing up in the case of the compound pulley (Mechanics 2.24, Fig. 12; since the Ms figure is not very clear, the following refers to the figure in Heronis Alexandrini opera). Specifically, Heron notes that in order to lift a weight of 1000 talents at over the distance , a force of 200 talents exerted at must be pulled through 5 times the distance (since each rope in the 5-pulley system must be pulled over a distance ). Similarly, a force of 40 talents exerted at must be pulled through 5 times the distance of the force at (since in order to move the rope at by a certain distance, the rope at must be pulled through 5 times that distance)",
+            ", R\u2032 F1 F2 D2 D1 where D1 and D2 are the distances covered by the moving forces, travelling at a constant speed, in different times.49 Now the operation of the compound pulley can also be analyzed in a way that leads to a different understanding of the phenomenon of slowing up. On this view, the comparison is between the moving force and the weight moved, rather than between two different moving forces. If the moving force is smaller than the weight, it will cover more distance than the weight in any given time (e.g., in Fig. 12 during the time in which the force at moves a certain distance, the weight at will move 1/5 of that distance). Thus the moving force travels more quickly than the weight; in this sense, the \u201cslowing up\u201d that occurs in the machine is the result of the weight moving more slowly than the force that moves it. At any given time the ratio of the distance covered by the moving force to that covered by the weight will be the inverse of the ratio of the force to the weight, i.e., R\u2032\u2032 F1 F2 D2 D1 where F2 equals the weight"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003636_s00542-012-1544-7-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003636_s00542-012-1544-7-Figure11-1.png",
+        "caption": "Fig. 11 Models with and without spiral grooves on thrust bearing",
+        "texts": [
+            " The air\u2013oil interface of the model with tapering angle 5 doesn\u2019t break up at 16,200 rpm, but it breaks up at 18,000 rpm. The air\u2013oil interface of the model with tapering angle 10 doesn\u2019t break up at 18,000 rpm, but it breaks up at 19,800 rpm. Table 2 shows the simulated break-up speed of air\u2013oil interface due to three different tapering angles. 4.2 Effect of the thrust grooves on the leakage of operating FDBs Oil leakage was investigated due to the inward pumping effect of spiral groove on the thrust bearing. Figure 11 shows the geometries of the finite volume models with and without spiral grooves on the thrust bearing. Figures 12 and 13 show the simulated unsteady motions of fluid lubricant and air\u2013oil interface due to the operating speed at 16,200 and 18,000 rpm. The air\u2013oil interfaces of both models do not break up at 16,200 rpm or less. The air\u2013oil interfaces of both models break up at 18,000 rpm and over. The oil moves downward along the hub and moves upward along the sleeve due to the centrifugal effect of oil"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003790_20110828-6-it-1002.01626-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003790_20110828-6-it-1002.01626-Figure5-1.png",
+        "caption": "Fig. 5. Virtual target.",
+        "texts": [
+            " The desired arrival specifications to the target located at position ntg are defined by the following vector: atg = [tarr Sarr] (2) where tarr is the desired arrival time to target (in number of sampling periods) and Sarr is the desired arrival speed vector (speed in x, y and z), thus defining the desired arrival path angle and heading. To respect the arrival specifications, it is proposed (Be\u0301langer et al., 2008) to create a virtual target that moves at the desired speed along a straight line intercepting the real target position at the desired arrival time (Fig. 5). The position of the virtual target at time k, ntgv(k), is therefore: ntgv(k) = ntg \u2212 ts \u00b7 (tarr \u2212 k) \u00b7 Sarr (3) where ts is the sample time. If the tracking of the virtual target by the UAV is successful, the arrival specifications will be respected. In the context of predictive control, at time k, it will be required to predict the virtual target trajectory over the prediction horizon, yielding ntgv(k+1:k+hp|k). These position predictions are simply obtained by moving the virtual target at constant speed in a straight line (Fig. 5). At each sample time k, the autopilot transmits the measured UAV position nm uv(k) to its TCU (Fig. 1, Fig. 4). Using nm uv(k), the TCU then employs an augmented extended Kalman filter (EKFuv), as detailed in Pre\u0301vost et al. (2009), to predict the optimal UAV trajectory n\u0302uv(k+1:k+hp|k) which accounts for both parametric model errors and unmeasured disturbances \u03be(k). At each sample time k, the TCU calculates the UAV setpoint trajectory over the control horizon u\u0304uv(k+1:k+hc|k), and consequently the optimal predicted UAV trajectory over the prediction horizon \u00af\u0302nuv(k+1:k+hp|k), that minimizes an objective function j(k) while respecting a set of linear and nonlinear constraints",
+            " Probability p(ei,\u03c4 ) is calculated by convolving the probability density function of the obstacle uncertain position (nobi (k+\u03c4 |k) in (1)) with an indicator function defining the obstacle shape (ellipsoidal of size qm obi (k) and attitude rm obi (k)). Probability theory is then used to calculate p(e\u03c4 ) as a function of all p(ei,\u03c4 ). Details on this calculation can be found in Pre\u0301vost et al. (2011). The target tracking criterion jtg(k) = hp \u2211 \u03c4=1 \u03batg(\u03c4) \u2016 n\u0302uv(k+\u03c4 |k) \u2212 ntgv(k+\u03c4 |k) \u2016 2 (7) seeks to limit the 2-norm distance \u2016 \u2022 \u2016 2 between the predicted UAV position and predicted virtual target position along hp, as illustrated in Fig. 5. Minimizing this criterion seeks to drive the UAV to intercept the target at the specified arrival time and with the specified speed and direction. In (7), the constant \u03batg(\u03c4) weights the UAVtarget distance along the trajectory. Numerical minimization of j(k) is subject to constraints. Firstly, the UAV setpoint increments on hc and UAV setpoints on hc must adhere to the UAV maneuverability constraints. Thus, \u2206umin uv \u2264 \u2206uuv(k+1:k+hc|k) \u2264 \u2206umax uv (8) umin uv \u2264 uuv(k+1:k+hc|k) \u2264 umax uv (9) Secondly, the predicted optimal UAV trajectory should be contained within the span of the aircraft sensor range hsn n\u0302uv(k+1:k+hp|k) \u2286 hsn (10) thus ensuring that a predicted UAV trajectory does not intersect an undetected obstacle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003995_j.procs.2011.08.077-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003995_j.procs.2011.08.077-Figure1-1.png",
+        "caption": "Figure 1: Kinematic configuration of accelerometer measurements in the inertial frame.",
+        "texts": [
+            " Section 2 presents the basic relationship between the accelerometer measurements and the vehicle acceleration; this relationship is utilized by Section 3 to lay out a comprehensive computational algorithm for this problem. The simulation results of the proposed algorithm are presented in Section 4 before closing the paper with some concluding remarks in Section 5. Assume that the frame xyz is attached to, and moves with the vehicle at its center of mass, G, such that the y axis is always pointing in the direction of the vehicle in the XYZ inertial frame. If a three axis accelerometer equipped with with a three axis gyro is attached to a vehicle as shown in Figure 1, then the accelerometer measurements \u0308rxyz do not directly represent the actual acceleration of the vehicle \u0308rXYZ in the inertial frame as required by the odometry system. It can easily be shown that that these quantities are related through \u0308rxyz = J\u0308rXYZ + 2J\u0307\u0307rXYZ + J\u0308 rXYZ (1) where matrix J is the X-Y-Z rotation matrix defined using standard Euler angles \u03b8x, \u03b8y, and \u03b8z as [12, 13] J = \u23a1\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 cos \u03b8y cos \u03b8z \u2212 sin \u03b8z cos \u03b8x + sin \u03b8x sin \u03b8y cos \u03b8z sin \u03b8x sin \u03b8z + cos \u03b8x sin \u03b8y cos \u03b8z cos \u03b8y sin \u03b8z cos \u03b8z cos \u03b8x + sin \u03b8x sin \u03b8y sin \u03b8z \u2212 sin \u03b8x cos \u03b8z + cos \u03b8x sin \u03b8y sin \u03b8z \u2212 sin \u03b8y cos \u03b8y sin \u03b8x cos \u03b8x cos \u03b8y \u23a4\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 (2) The accelerometer measurements include the Coriolis component, 2J\u0307\u0307rXYZ , and the rotational component , J\u0308 rXYZ , which are not required by the odometry system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002285_660432-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002285_660432-Figure4-1.png",
+        "caption": "Fig. 4 - View of oscillating arm, relative rotation machine",
+        "texts": [],
+        "surrounding_texts": [
+            "2 the bolt be caused to loosen by the process of fluctuating tension alone. This experiment was conducted under various values of fluctuating tension, ranging from extremely small\nvalues up to large values, both applied over long periods of time. In no case was bolt rotation observed. This conclusion must be specifically limited by noting that it applies only to a bolt tightly torque into a threaded hole. It does not apply to a nut-bolt system such as that used by Goodier and Sweeney.\nIt is entirely possible that by the use of a nut, extremely small nut rotations could be forced by large fluctuating tensions. However, since the primary interest here is self-lock -\ning bolts, it must be reported that no success has been obtained with the process of fluctuating tension, which so often is pointed out in theoretical machine design as the cause of bolt loosening.\nWhile it is true that, in many machine situations, fluctu-\nating tension remains the primary component of load on a bolt, there are also many situations in which two mating parts held by a bolt rotate a minute amount with respect to one another. Under such conditions the bolt may be observed to loosen and eventually come out. This can be demonstrated in the laboratory in at least two specific ways, both of which utilize the basic phenomena illustrated in Fig. 1. Here, one may see that two primary elements are needed in order to cause loosening of a bolt by oscillating torque or rotation. The first of these is the stationary portion, marked A in Fig. 1, composed of the base rigidly attached to some massive foundation. To this base is attached an ear E having a body hole so that the bolt B may be put through the hole and screwed tightly into the part C. This simulates the attachment of one part to another by a bolt, in the usual way.\nAt this point two possible mechanisms of loosening are\navailable. In the first, described in Ref. 1, a small fluctuating angular displacement is imposed on the part C, generally of harmonic form A e sin wt, and the resulting motion of bolt B is observed. In some cases,.secondary effects tend to cloud the issues somewhat, but in general the results of observations on bolt B are given schematically in Fig. 2, where the curve is drawn through points at which the bolt loosens.\nrotation\nHere, one sees that above a certain value of angular amplitude A e, the bolt always loosens; in general, at higher values of A 0, this loosening is extremely rapid.\nBelow this limiting value of 06, it appears from experi-\nmental evidence presently available that no amount of cyclic oscillation will result in loosening of the bolt. In some respects, this curve is similar to a fatigue curve that defines a stress level at which a material will operate indefinitely. In this exploratory case, one may define by analogy an oscillatory amplitude level that is safe for design purposes. To the best of the writer's knowledge, this phenomenon was first described in Ref. 1.\nDESCRIPTION OF EQUIPMENT\nSince a typical machine element in which a self-locking\nbolt is used is not generally subjected to tension conditions alone, it seems quite possible that small relative rotations between parts may in many cases be responsible for the loosening of bolted connections.\nIn order to define clearly the effects of relative rotation, a machine was designed and constructed similar in principle to that shown in Fig. 1, but with some added instrumentation and measuring devices. A photograph of this machine is shown in Fig. 3; Figs. 4 and 5 show detailed views of the bolt B, ear E, and head section C of the machine. Fig. 5 also illustrates the measuring device used to indicate looseness of the bolt.\nIn the design of this machine, it should be noted that an\ninternal strain gage dynamometer is built into the ear. This dynamometer is used for measuring axial load in the bolt as it is tightened into the head section C. This allows each test bolt to be torqued to some consistent value of axial load. This conforms with practices associated with static testing, for reasons discussed in Ref. 1.\nWhen tightening a bolt into the fixture, output of the\nstrain gage dynamometer indicates the bolt axial force. When the proper force has been reached, the motor is switched on and the arm oscillates through some small angle (which can be preset before each experiment) at a fixed frequency of 360 cpm. Dynamometer readings indicate that bolt tension fluctuates over a small range. However, the primary effect seems to be relative rotation of the ear E and head C of the fixture.\nCYCLES TO LOOSEN\nrig. 2 - Typical data showing angular, eccentricity versus\ncycles to loosen\nLOCUS OF POINTS AT WHICH BOLTS LOOSEN",
+            "3 If the oscillatory angle of the arm is large enough, bolt tension drops off rapidly as cyclic relative rotation occurs. The bolt \"rachets\" loose. This is observed by monitoring the output signal from the strain gage dynamometer under the bolt head. This slow \"ratcheting\" continues until the test is stopped or until the bolt has lost all or most of its tension in the shank.\nIn either event, the number of cycles can be counted and recorded. On the other hand, if the oscillatory angle of the arm is small enough, there is apparently no tendency of the bolt to ratchet loose, and a large number of cycles of motion have been observed to cause no reduction in bolt tension. These results are compatible with the curves shown in Fig. 2.\nOne of the primary difficulties associated with such a machine is the unavoidable elastic deformation occurring at each stroke in the ear and arm system. As was previously noted, the original eccentricity of the machine may be preset by means of adjusting the throw of the eccentric. However, the large forces carried here cause elastic deformations, which are of the same order of magnitude as the preset eccentricities when these eccentricities are small, so that in this case the actual values of D e are very difficult to measure. Under such conditions, considerable doubt exists as to the validity of the data in this region, and primarily for this reason one finds difficulty in fully accepting the results of tests from this machine at low eccentricity values.\nIn order to avoid this difficulty of unknown total eccentricity, it was decided to construct a machine operating on principles similar to that shown in Fig. 1, but now modified in such a way that a fluctuating torque would be applied to\nthe member C of Fig. 1. If such a fluctuating torque could be applied independently of the angle of rotation of part C, then all doubt would be removed concerning the exact variables used in a particular test, The fluctuating torque chosen at the beginning of the test should remain constant throughout the duration of the test, independent of the magnitude of the resulting oscillatory motion. In order to accomplish this, it was necessary to utilize some sort of device that was capable of exerting forces independent of oscillatory amplitude.\nExamination of commercially available equipment indicated one particular machine well suited for this type of operation, a machine specifically designed for fatigue studies in such a way that the force amplitudes are independent of the displacement of the specimen. This particular machine operates on a resonant force principle, and is sold under the trade name \"Baldwin-Sonntag\" fatigue testing machine. It may be obtained through the various sales offices of the Baldwin-Lima-Hamilton Corp. Since this particular type of fatigue machine is rather well known for testing work, no effort will be made here to discuss the details of its dynamic operation. However, a brief description of its primary parts is probably in order.\nIn this machine, oscillatory forces are produced by a rotating counterweight driven at 1800 rpm by a synchronous motor. The eccentricity of the rotating counterweight may be varied in order to apply different values of fluctuating force to a specimen. The forcing frequency is, of course, fixed at 1800 rpm or 30 cps. The entire moving portion of the machine, along with the appropriate portions of the specimen fixtures, are \"tuned\" into a resonant condition, which",
+            "4\ncauses the force applied to the specimen to be essentially independent of displacement. As a crack progresses in a typical fatigue specimen and amplitude of displacement increases, the forces will indeed remain constant until fracture is complete. Such machines are typically equipped with a microswitch displacement shutoff control so that at large displacements, coinciding with the fracture of a specimen, the machine automatically shuts off.\nIn this particular instance, a 100 lb capacity Baldwin -Sonn-\ntag fatigue testing machine was equipped with its standard torsion fixtures so modified as to produce the type of action illustrated in Fig. 1, except that now part C of Fig. 1 was subjected to an oscillatory torque AT sin cot rather than an oscillatory displacement.\nA photograph of the modified fixture is shown in Fig. 6,\nand the torque or moment arm is shown in the background of Fig. 7. From these photographs it may be seen that the small platen which moves up and down in an oscillatory fashion, transmitting the unbalanced eccentric force, is coupled by means of a torque arm directly through a part similar to arm C of Fig. 1. This, in turn, is held to a flat plate by means of the test bolt. The flat plate is securely fixed to the base of the testing machine by means of two heavy supports, as shown in Fig. 7.\nFig. 8 shows an element of the test fixture corresponding\nto part C of Fig. 1. This rod is, of course, threaded to receive the test bolt and is further specially constructed to measure the axial load in that bolt. The form of this construction is shown in Fig. 9, where it is seen that a portion of the cylindrical member is removed so that a thin cylindrical shell can be substituted. This shell has been instrumented with strain gages so that axial load may be accu-\nrately measured.\nTESTING PROCEDURE\nThe procedure for testing a bolt under a particular set of\nconditions begins by inserting the bolt through the fixed mem - ber and into the threaded cylindrical portion, followed by tightening the bolt to some level of shank stress as indicated by the strain gage instrumented dynamometer on the shoulder of the cylindrical portion. This is continuously monitored\nby a strain gage bridge.\nAfter obtaining the desired bolt tension, the eccentric\nweight of the machine is adjusted to the value desired for this test. The machine is then started rapidly and the oscillating torque caused by the oscillating force is applied\nat the end of the moment arm and allowed to act on the bolt. If the eccentric force is too large, the bolt will come loose\neter, relative rotation machine\ntorque on bolted parts"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000827_978-0-585-34652-6_10-Figure18-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000827_978-0-585-34652-6_10-Figure18-1.png",
+        "caption": "Figure 18 The structure of two-joint manipulator",
+        "texts": [
+            " Tracking Now, suppose that we want to control the states of the pole to track a desired sinu soidal trajectory given by x,j = a sin(M) x^j = dx,j / dt = aco cos(CiX) Then we can select the sliding surface as <y= ix,-Xij) + (x2-xJ Again, adopting the FSMC as given in Example 1, we get the system response as shown in Figure 15. Similarly, by adding the load variation and external disturbance shown in Fig ure 13b and Figure 14b, we get the resulting trajectories shown separately in Figure 16 and Figure 17. 2. Two-Join Manipulator Consider a two-joint manipulator shown in Figure 18, where g is the gravitational con stant, 0, and 0j are joint angles, and u, and M\u0302 are the applied torques. Let Xi =6 i , X2 =6] , X3 =02, ^4 =02 and under the assumption that there exist lumped equivalent masses and massless links, then its dynamics can be represented as the follow ing state space [20][22]: Xl ~ Xy Xl = -f(^,2>c^' + 2P,2A:2A:, + y,g + M,) - ^ ( - p . ^ x , ' + y,g + \u00ab,) ^ m v2.i.oft . . ^ \u201e \u201e ^ . . ^ M / f t . 2 . X3 \u2014X^ X4 = - --f-(Pi2^2' + 2p,jX^, + Y,g + M,) + -7-(-p,iX,' -I- Yzg + \u00ab,) with a,, = (m, + w^) / / -t- 7M\u0302 /j^ + Im^ I, I2 cos Xj -H J, a,2 \u2014m^ I2 + w^ /, I2 cos Xj Y, = -{{m, + m2)lj cos x^ + m212 cos(X; -I- Xj)) Y2 = - W j /j COS(X; + Xj) 301 302 303 Px2 = Wi 11 h sin X, where m,, I"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000058_0020-7403(80)90031-4-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000058_0020-7403(80)90031-4-Figure12-1.png",
+        "caption": "FIG. 12. Effect of varying the gravity parameter W..",
+        "texts": [
+            " The variation of the maximum force P transmitted to the bearing housing with the frequency ratio f / i s shown on Figs. 9 and 10. Fig. 9 is for the case where U, s r and B are kept constant while c~ and W are varied, whereas Fig. l0 is for the case where U, s r and a are kept constant while B and ta/are varied. Fig. l l is again a plot of the maximum orbit amplitude ratio G vs the frequency ratio 1~. However, in this instance the exercise is to illustrate only the effect of varying if\" while B, U, ~\" and c~ are kept constant. Fig. 12 is also for the case where B, U, r and a are kept at constant realistic values, but in this instance shows the rotor centre and journal centre orbits at the first pin-pin resonance of the system for various values of the gravity parameter W. DISCUSSION The set of rotor orbits shown in Fig. 5, has an interesting trend (the journal orbits tend to be too small for trends to be detected). They begin with extremely small elliptical excursions whose axes have positive slope, and then as the excursions grow in magnitude, the major axis precesses in the direction of rotation",
+            " 6, or it may have an even more irregular shape as is the case for the B = 0.05, W = 0-10 condition shown in Fig. 8. These deviations from the single peak response are presumably due to the non-linear asymmetric characteristics of the fluid-film and the simultaneous effect of both the gravity loading and mass unbalance. When the value of the gravity parameter v~, is reduced to zero (equivalent to a vertically mounted rotor shaft), the rotor orbits obtained from the present investigation are always circular and centred about the bearing centre. As can be seen from Fig. 12, the overall effect of a reduction in l~' is to reduce the working eccentricity and the magnitude of the major axis of the elliptical orbit until, when I$' = 0, the orbits become circular and centred. The influence of the gravity parameter W on the excursion amplitude ratio G is shown in Fig. ll . From this figure it can be seen that the response of the system approaches the I~ = 0 case asymptotically for values of the frequency ratio fl, which are near to the rigid support resonance (fl = 1.0) or greater"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000760_robot.2007.363072-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000760_robot.2007.363072-Figure1-1.png",
+        "caption": "Fig. 1. Reference frames",
+        "texts": [
+            " Stability analysis of the closed loop system is given in section 5 and simulations results are provided in section 6. Concluding remarks are finally given in the last part of this paper. A. VTOL UAV model The VTOL UAV model is represented by a rigid body of mass m and of tensor of inertia I = diag(I1, I2, I3) with I1, I2 and I3 strictly positive. We define an inertial reference frame (I) associated with the vector basis (e1, e2, e3) and a body frame (B) attached to the UAV and associated with the vector basis (eb1, e b 2, e b 3) (see Fig. 1). The position and the linear velocity of the UAV in (I) are respectively denoted 1-4244-0602-1/07/$20.00 \u00a92007 IEEE. 724 \u03be = [x y z] T and v = [vx vy vz] T . The orientation of the UAV is given by the orientation matrix R \u2208 SO(3) from (I) to (B), usually parameterized by Euler\u2019s pseudo angles \u03c8, \u03b8, \u03c6 (yaw, pitch, roll): R =   c\u03b8 c\u03c8 s\u03c6 s\u03b8 c\u03c8 \u2212 c\u03c6 s\u03c8 c\u03c6 s\u03b8 c\u03c8 + s\u03c6 s\u03c8 c\u03b8 s\u03c8 s\u03c6 s\u03b8 s\u03c8 + c\u03c6 c\u03c8 c\u03c6 s\u03b8 s\u03c8 \u2212 s\u03c6 c\u03c8 \u2212s\u03b8 s\u03c6 c\u03b8 c\u03c6 c\u03b8   (1) with the trigonometric shorthand notations c\u03b1 = cos(\u03b1) and s\u03b1 = sin(\u03b1), \u2200\u03b1 \u2208 R"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001230_12.827720-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001230_12.827720-Figure2-1.png",
+        "caption": "Fig. 2. (a) Schematic of process flow, (b) photograph of device active area, and (c) photograph of device with test contact pads.",
+        "texts": [
+            " Immobilization of GOx onto ITO is formed by stable ester bonds resulting from -OH groups on the ITO surface and the carboxylic groups on the enzyme shell [9,10]. The action of the enzyme on glucose generates hydrogen peroxide that is detected electrochemically using the oxidation current on the platinum electrodes, which are Proc. of SPIE Vol. 7397 73970K-2 Downloaded From: http://ebooks.spiedigitallibrary.org/ on 10/25/2013 Terms of Use: http://spiedl.org/terms referenced to an external silver/silver chloride electrode. GOx immobilization allows for precise quantification of the amount of enzyme near the platinum electrode The fabricated design (Fig. 2) has two devices and an on-chip reference electrode, however the results we present do not make use of the on-chip reference electrode. The left device does not have an ITO region and is used as a control device to check the quality of the enzyme immobilization. The working electrode is a pattern of 50 \u00b5m wide incomplete concentric rings that are connected by a common radius. The counter electrode is a 75 \u00b5m wide ring with an overall surface area greater than the working electrode so that the current signal is a function of the area of the working electrode only"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003749_s1068366612020110-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003749_s1068366612020110-Figure4-1.png",
+        "caption": "Fig. 4. Schematic of measuring friction torque on SI 03M test machine: ICS\u2014information and control system; PC\u2014personal computer.",
+        "texts": [
+            " For this purpose the friction torque meter is mounted directly on the motor shaft of the drive of the roller counterspecimen; this makes it possible to avoid bearing loss. The torque is measured by a precision T10F gage (Hottingen Baldwin Messtechnik). This gage is used for measuring dynamic torques on rotating shafts and does not include bearings and coll rings. As a result, frictional and heat effects in the bearings are eliminated. The accuracy of the measurement of the friction torque by the gage is 0.1%. Figure 4 shows the schematic of mea suring the friction torque on the SI 03M test machine. The coefficient of resistance to rolling during the tests is calculated as follows: (1) where Mfr is the measured friction torque, N m; FN is the contact load, N; Rc = 0.05 m is the radius of the roller counterspecimen. The tests on the SI 03 machine involve the full automatization of the measurement and recording of the characteristics under study. The following data are displayed on the monitor screen of the control com puter: the contact load, the bending load, the approach of the axes, the friction torque, the friction temperature, the rotational velocities of the specimen and counterspecimen, the calculated slippage coeffi fr r c , N M f F R = Characteristics of materials under testing Material Heat treatment Surface hardness Endurance limit at Torsional Bending \u03c3\u20131, MPa Tensile yield point \u03c3y, MPa* Tensile strength \u03c3t, MPa* Steel 25KhGT Surface hardening 700 HV 760 900 1700 Steel 45 Normalizing 470 HV 270 380 610 Note"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003709_012082-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003709_012082-Figure7-1.png",
+        "caption": "Figure 7. FE model of gearbox casing.",
+        "texts": [
+            " Note that the low order modes are greatly affected by the restraints applied to the master DOFs, but boundary conditions have little effect on the frequency of the high order modes defining the frequency range. The decomposition of the model into both physical DOFs (master DOFs) and modal coordinates allows the flexibility of connecting the finite elements to other substructures, while maintaining a reasonably good result within a specified frequency range. Our earlier studies had shown that the simple LPM gave good results at high frequencies, and the interaction problems were primarily in the mid frequency range. The FE model of the gearbox casing (with 104 340 DOFs) is shown in figure 7 and was modelled using both solid and shell elements. The support of the casing by rubber pads was simulated using spring elements at the corners of the casing. The model has been compared with experimental modal testing and validated for lower frequency modes [9]. The centre node of each bearing was selected as a master degree of freedom which enables connecting the reduced model of the casing with the LPM model of the internals, thereby capturing the flexibility of the casing. The reduced mass and stiffness matrices of the casing (obtained using the Craig-Bampton based CMS reduction method) were combined with the LPM model in the Simulink environment"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002930_00207179.2011.625044-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002930_00207179.2011.625044-Figure1-1.png",
+        "caption": "Figure 1. The information graphs: (a) G(L1), (b) G(L2) and (c) G(L3).",
+        "texts": [
+            " D ow nl oa de d by [ U ni ve rs ity o f C hi ca go L ib ra ry ] at 0 2: 31 2 6 D ec em be r 20 14 To illustrate the developed results, two simulation examples on a directed network of four nodes are provided for the periodic switching and randomly switching, respectively. The nodal self dynamics is the well-known Lorenz system illustrated in the following: f \u00f0xi\u00de \u00bc 10xi1 \u00fe 10xi2 28xi1 xi1xi3 xi2 xi1xi2 8 3 xi3 0 BBB@ 1 CCCA, with xi \u00bc xi1 xi2 xi3 0 BB@ 1 CCA: It is known (Christiansen and Rugh 1997) that the maximum Lyapunov exponent of the above Lorenz system is max\u00bc 0.9057. 5.1 Periodic switching In the first example, the switching signal (t) is periodic with the cycle time T\u00bc 3 s. The topology collection M\u00bc {1, 2, 3} are shown in Figure 1, whose elements correspond to the following Laplacian matrices, respectively, L1 \u00bc 3:5 3:5 0 0 0 3:5 3:5 0 0 0 0 0 0 0 0 0 2 6664 3 7775, L2 \u00bc 2:5 0 2:5 0 0 0 0 0 0 0 0 0 2:5 0 0 2:5 2 6664 3 7775, L3 \u00bc 4 0 0 4 0 0 0 0 0 4 4 0 4 0 0 4 2 6664 3 7775: Every second, (t) updates its value by order of 1, 2, 3, 1, 2, . . . and correspondingly the topology updates by order of L1, L2, L3, L1, L2, . . . . Without loss of generality, assume that (0)\u00bc 1. It should be emphasised that the networked system over any one single graph among the collection of {G(Li)} will not be synchronous because there always exists the isolated node in Li for all i, as shown in Figure 1. While considering the transition matrix (kT\u00fe t, (k 1)T\u00fe t) with t\u00bc 0 and T\u00bc 3, \u00f0kT, \u00f0k 1\u00deT \u00de \u00bc eL 3 eL 2 eL 1 \u00bc 0:0043 0:0152 0:9395 0:0410 0 0:0302 0:9698 0 0 0:0296 0:9704 0 0:0043 0:0152 0:9394 0:0411 2 6664 3 7775\u00bc4 S: \u00f05:1\u00de Its contract rate is \u00f0S\u00de \u00bc kQTSQk \u00bc 0:0535, \u00f05:2\u00de where Q2RN (N 1) is an orthogonal basic matrix of space orthogonal to span(b). For periodic switching, this value is also the consensus convergent rate (3, 0) of digraph. Noting that the union of G(L1), G(L2) and G(L3) has a spanning tree embedded it, Assumption 1 holds here for T\u00bc 3 and t\u00bc 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001121_j.na.2007.02.015-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001121_j.na.2007.02.015-Figure2-1.png",
+        "caption": "Fig. 2. Orbits of the unperturbed system.",
+        "texts": [
+            "1) y 7\u2192 \u2212(1 \u2212 \u03b5\u03c1)y as |x | = 1. (2.2) For \u03b5 = 0, the unperturbed system of (2.1) and (2.2) describes the following free impact oscillator: x\u0307 = y, y\u0307 = \u2212g(x), |x | < 1, (2.3) y 7\u2192 \u2212y, |x | = 1. (2.4) System (2.3) is a planar Hamiltonian system with Hamiltonian H(x, y) = y2/2 + G(x), where G(x) := \u222b x 0 g(s)ds, and by Hypothesis (H1) has a saddle at the origin O(0, 0), which is the unique equilibrium for x \u2208 (\u22121, 1). The phase portrait of system (2.3) and (2.4) is topologically equivalent to that shown in Fig. 2. Besides, the unperturbed system (2.3) and (2.4) has two homoclinic loops \u0393+ = O AA\u2032O and \u0393\u2212 = O B B \u2032O via the identification given by the impact law (2.4), where A\u2032, B \u2032 are the reflection points of A, B respectively. Both \u0393+ and \u0393\u2212 correspond to the level curve H(x, y) = 0. There are two families \u0393\u00b1 h of periodic orbits inside the homoclinic loops with periods ranging from zero (at the wall x = \u00b11) to the limit \u221e (along \u0393\u00b1), which correspond to the level curves H(x, y) = h parameterized by h \u2208 (G(1), 0) \u222a (G(\u22121), 0), where G(1) < 0 and G(\u22121) < 0 by (H1)",
+            " Then x(\u03b8m + mT ; \u22121,\u2212(1 \u2212 \u03b5\u03c1)yc,m,\u03b5, \u03b8m + tc) = 1, y(\u03b8m + mT ; \u22121,\u2212(1 \u2212 \u03b5\u03c1)yc,m,\u03b5, \u03b8m + tc) = ym,\u2223\u2223x(\u03b8m + t; \u22121,\u2212(1 \u2212 \u03b5\u03c1)yc,m,\u03b5, \u03b8m + tc) \u2223\u2223 < 1, t \u2208 (tc,mT ). Namely, the pendulum starting from the wall x = +1 with the initial velocity ym at the initial time \u03b8m immediately changes its velocity to \u2212(1\u2212\u03b5\u03c1)ym by the impact law and reaches the wall x = \u22121 after some time tc (0 < tc < mT ), and then bounces back to the wall x = +1 with the same velocity ym after travelling for another time of mT \u2212 tc without touching the walls x = \u00b11. From Fig. 2, it is easy to see that for each h \u2208 (0,+\u221e), the periodic orbit \u0393 0 h of the unperturbed system (2.3) and (2.4) corresponding to the level curve H(x, y) = h has two branches symmetric with respect to the x-axis: one is in the upper half-plane and is denoted by \u0393 0,+ h ; the other is in the lower half-plane and is denoted by \u0393 0,\u2212 h . By assumption (H3), for any given integer m \u2265 1, the unperturbed system (2.3) and (2.4) has a unique resonant periodic orbit \u0393 0 hm of period T 0 hm = mT such that its upper (lower) branch \u0393 0,+ hm : (xhm 0,+(t \u2212 t0), yhm 0,+(t \u2212 t0)) (resp",
+            " (3.1) Substitute (3.1) into (2.1); we find that (x h\u0304 1,\u2212(t, \u03b8\u0304 ), yh\u0304 1,\u2212(t, \u03b8\u0304 )) is the solution of the following linear variational system:{ x\u0307 h\u0304 1,\u2212(t, \u03b8\u0304 ) = yh\u0304 1,\u2212(t, \u03b8\u0304 ), y\u0307h\u0304 1,\u2212(t, \u03b8\u0304 ) = \u2212g\u2032(x h\u0304 0,\u2212(t \u2212 \u03b8\u0304 ))x h\u0304 1,\u2212(t, \u03b8\u0304 )+ f (t, x h\u0304 0,\u2212(t \u2212 \u03b8\u0304 ), yh\u0304 0,\u2212(t \u2212 \u03b8\u0304 )) (3.2) associated with the initial condition (x h\u0304 1,\u2212(\u03b8\u0304 , \u03b8\u0304 ), yh\u0304 1,\u2212(\u03b8\u0304 , \u03b8\u0304 )) = (0, \u03c1 y\u0304). Since the phase portrait of the unperturbed system (2.3) and (2.4) is symmetric with respect to the x-axis as seen in Fig. 2, it is easy to see that the time taken by the pendulum along the orbit \u0393 0,\u2212 h\u0304 from x = +1 to x = \u22121 is T 0 h\u0304 /2. Thus the orbit \u0393 0,\u2212 h\u0304 reaches the wall x = \u22121 at the time \u03b8\u0304 + T 0 h\u0304 /2. By the continuous dependency on parameters and initial values, when |\u03b5| > 0 is sufficiently small, the trajectory \u0393 0,\u2212 \u03b5,h\u0304 reaches \u03a3\u2212 \u2212 at some point P0 \u03b5,\u2212(\u03b8\u0304) at some finite time T 0 \u03b5,\u2212(\u03b8\u0304) > \u03b8\u0304 . Clearly, the orbit \u0393 0,\u2212 h\u0304 does not have a contact point with the walls x = \u00b11 until it reaches x = \u22121, and along the orbit \u0393 0,\u2212 h\u0304 , the impact velocity is not zero",
+            " Application to a nonlinear system In what follows we continue the work of [10,9] to discuss Type II subharmonic orbits of the nonlinear impact system whose free flight motion between the walls is given by the Duffing equation x\u0308 + \u03b5\u03b4 x\u0307 \u2212 x + x3 = \u03b5\u03b3 cos\u03c9t, for |x | < 1, associated with the impact law given by the second relation in (1.1). The system is equivalent to x\u0307 = y, y\u0307 = x \u2212 x3 + \u03b5(\u2212\u03b4y + \u03b3 cos\u03c9t) as |x | < 1, (6.1) y 7\u2192 \u2212(1 \u2212 \u03b5\u03c1)y as |x | = 1. (6.2) The unperturbed system of (6.1) is a planar Hamiltonian system with Hamiltonian H(x, y) = 1 2 y2 \u2212 1 2 x2 + 1 4 x4. When \u03b5 = 0, the phase portrait of (6.1) and (6.2) is topologically equivalent to that shown in Fig. 2. Here T = 2\u03c0/\u03c9. By the result in [22, p.154], outside the homoclinic orbits the unperturbed system (i.e. \u03b5 = 0) has the upper branch \u0393 0,+ h : (xhk 0,+(t), yhk 0,+(t)) of the family of periodic orbits based on x = \u22121 and the lower branch \u0393 0,\u2212 h : (xhk 0,\u2212(t), yhk 0,\u2212(t)) of the family of periodic orbits based on x = +1, i.e., (xhk 0,+(t), yhk 0,+(t)) = ( \u221a 2k \u221a 2k2 \u2212 1 cn (\u03c8k(t), k) , \u2212 \u221a 2k 2k2 \u2212 1 sn (\u03c8k(t), k) dn (\u03c8k(t), k) ) , (xhk 0,\u2212(t), yhk 0,\u2212(t)) = ( \u221a 2k \u221a 2k2 \u2212 1 cn (\u03c6k(t), k) , \u2212 \u221a 2k 2k2 \u2212 1 sn (\u03c6k(t), k) dn (\u03c6k(t), k) ) for t \u2208 [0, T 0 hk /2], where \u03c8k(t) = t \u221a 2k2 \u2212 1 + 2K (k)+ F (\u03d10, k) , \u03c6k(t) = \u03c8k(t)\u2212 2K (k), T 0 hk = 4 (K (k)\u2212 F (\u03d10, k)) \u221a 2k2 \u2212 1, \u03d10 := arccos (\u221a 1 \u2212 1 2k2 ) , sn, cn and dn are the Jacobi elliptic functions, K (k) is the complete elliptic integral of the first kind, F(\u03d5, k) is the incomplete elliptic integral of the first kind (see [4, p"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001802_cae.20219-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001802_cae.20219-Figure5-1.png",
+        "caption": "Figure 5 CAD models and its virtual objects of automotive chassis. [Color figure can be viewed in the online issue, which is available at wileyonlinelibrary.com.]",
+        "texts": [
+            " The latest version of VRML (e.g., VRML 2.0) included a special interface called External Authoring Interface (EAI) [35] which adds significant functionality to the language. This functionality concerns the handling of VRML scene from an external source such as a Java program and it can be used to add streaming capabilities to VRML [36 38]. The process of building virtual scenes in the virtual learning system of this study. First, CAD/CAM is used to construct the subsystems in the chassis. As shown in Figure 5a, the CAD models of transmission, steering and suspension subsystems. Then, the components of each subsystem are converted into STL files and imported the animation generation software for the editing of the models. The editing includes (1) triangular mesh optimization of the models in order to reduce file size and transmission time, (2) defining the absolute coordinate origin and relative position of each component in order to facilitate the designing of exploding/assembly paths and (3) exporting wrl files (VRML format) to form virtual models as shown in Figure 5b. In addition, VrmlPad is used to edit the virtual action program codes and to define the object which users need to control and its positions. The positions of VRML objects are accessed and set through EAI library. In exporting scene graphs field, given a scene graph that can export it to a VRML browser that supports the EAI interface. For exporting a Java scene graph to a browser, a browser object has to be obtained and handled, they can be done by calling the appropriate methods of the EAI. Subsequently, to show the virtual objects in the browser, the following line of code is needed: browser"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001104_09544062jmes817-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001104_09544062jmes817-Figure3-1.png",
+        "caption": "Fig. 3 Forces and moments acting on a ball in XOZ plane",
+        "texts": [
+            " The contact forces between the rolling element and the rings are [1] Q1i = ( 2 \u03b4\u2217 1i )1.5 \u00b7 2E \u2032 1 3 (\u2211 \u03c11 )0.5 \u00b7 (\u03b41i) 1.5 + (19) Q2i = ( 2 \u03b4\u2217 2i )1.5 \u00b7 2E \u2032 2 3 (\u2211 \u03c12 )0.5 \u00b7 (\u03b42i) 1.5 + (20) where \u2018+\u2019 stands when the force is positive when \u03b41i 0, \u2211 \u03c11 and \u2211 \u03c12 are the curvature sum, and \u03b4\u2217 1i and \u03b4\u2217 2i may be obtained using the simplified solution in reference [1]. In order to simulate actual working conditions, centrifugal force, gyroscopic moment, inertia force, and friction are considered. Figure 3 shows the forces and moments acting on a ball in the X - and Z-directions. F1i and F2i are the friction of the ith ball, Fzi the centrifugal force of the ith ball, Myi the gyroscopic moment of the ith ball. According to Newton\u2019s theorem, the equations of the ith ball can be represented as follows \u2212Q1i sin \u03b11i + Q2i sin \u03b12i \u2212 F1i cos \u03b11i + F2i cos \u03b12i = 0 (21) \u2212Q1i cos \u03b11i + Q2i cos \u03b12i + F1i sin \u03b11i \u2212 F2i sin \u03b12i + m\u03c92 oiri = 0 (22) If the bearing displacements are given, the contact forces could be computed from equations (1) to (22) using the Newton\u2013Raphson method"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001726_gt2010-23683-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001726_gt2010-23683-Figure1-1.png",
+        "caption": "Figure 1. Foil bearing schematic",
+        "texts": [
+            " NOMENCLATURE \u03c9 Frequency \u03a0i Non-dimensional group Ai Acceleration, frequency domain bi j Damping coefficient bu Damping associated with u Fi Force, frequency domain fi Force, time domain fu Force associated with u Guv Power spectral density of arbitrary functions u and v Hi j Frequency response function i Directional index (x corresponds to horizontal direction, y corresponds to vertical) j Directional index (x corresponds to horizontal direction, y corresponds to vertical) ki j Stiffness coefficient ku Stiffness associated with u mu Mass associated with u Si Positional coordinate, frequency domain si Positional coordinate, time domain t Time u Function index (placeholder) v Function index (placeholder) 1 1 Copyright \u00a9 2010 by Her Majesty the Queen in Right of Canada Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/gt2010/70410/ on 03/12/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Foil bearings are compliant-surface, hydrodynamic rotor supports, normally constructed from multiple layers of sheet metal foils. A generalized bearing, depicted in Figure 1, features a top-foil supported by one or more elastic sub-foils. The top-foil, which is typically coated with a wear-resistant, anti-friction material, traps an air film that supports a load. As the gas pressure on the top-foil changes, the force is transmitted to the elastic sub-foils, which vary by bearing design but are most often corrugated, spring-temper sheet metal. Deflection of the top-foil (under the influence of the gas film pressure) causes the sub-foils to compress radially and expand circumferentially"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001283_s11029-010-9111-8-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001283_s11029-010-9111-8-Figure5-1.png",
+        "caption": "Fig. 5. Re la tions be tween the load p0 and de flec tion u 3 M with ( ) and with out (- - -) ac count of the fol lower load for the rub ber-cord to roid al shell.",
+        "texts": [
+            "2 mm, and the ori en ta tions of rub ber-cord lay ers g gk k= - -( )1 1 , where g = 45\u00b0 and k = 1 4, . As a ref er ence sur face, we as sume the shell midsurface formed by ro ta tion of a part of the cir cum fer ence of ra dius R = 50 mm. The dis tance from the ro ta tion axis to the equa tor of midsurface R0 = 250 mm; the cross sec tions of the shell with the co or di nates \u00b1120\u00b0 are rigidly clamped. Cal cu la tion re sults for the tire un der the ac tion of fol lower and con ser va tive loads were ob tained by us ing reg u lar 24 \u0301 1 fi nite-el e ment meshes at e = 10\u20136 (see Ta ble 3 and Fig. 5). As is seen, the data in Ta bles 1 and 3 agree in a qual i ta tive sense with each other. In par tic u lar, the so lu tion of this prob lem for a di ag o nal tire can also be found with out us ing the in cre men tal ap proach. How ever, here the in flu ence of the fol lower load does not ap pear so noticeably (see Figs. 3 and 5). In the fol low ing in ves ti ga tion, the pres ent au thors are go ing to gen er al ize the re sults ob tained and use them for cal cu - lat ing a rub ber-cord shell of rev o lu tion in the pres ence of uni lat eral restrictions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-163-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-163-1.png",
+        "caption": "Figure 3-163: Vaporization shortens seal life and impairs seal performance.",
+        "texts": [
+            " \u2022 Check, and if necessary, improve the cooling and lubrication at the seal faces. \u2022 Consult knowledgeable manufacturers for recommendations on reducing seal generated heat. 2. Vaporization Symptoms and Failure Modes. Any popping, puffing, or blowing of vapors at the seal faces is evidence of vaporization. Vaporization does not frequently cause catastrophic failure, but it usually shortens seal life. Inspection of the seal faces reveals signs of chipping at the inside and outside diameters and pitting over the entire area. See Figure 3-163. Machinery Component Failure Analysis 237 Causes. Vaporization often occurs when heat generated at the seal faces is not adequately removed, and the liquid between the seal faces boils or flashes. Vaporization can also be caused by operating the seal too near the vapor pressure of the product in the seal cavity. Other operating conditions that will bring about vaporization include: \u2022 Excessive pressure for a given seal. \u2022 Excessive seal face deflection. \u2022 Inadequate cooling and lubrication of the seal"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001283_s11029-010-9111-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001283_s11029-010-9111-8-Figure1-1.png",
+        "caption": "Fig. 1. Ge om e try of a shell.",
+        "texts": [
+            " 7-Pa ram e ter Model of a Com pos ite Shell Let us con sider a thin shell of thick ness h d d= +- + con sist ing of NL elas tic anisotropic lay ers of con stant thick ness hk . We as sume that, at each point of the shell, there ex ists a sur face of elas tic sym me try par al lel to a ref er ence sur face W. As the ref er ence sur face, we con sider an in ner sur face of some k-layer or an in ter face be tween lay ers and re late it to curvilinear or - thogo nal co or di nates q1 and q2 , reck oned along the lines of main cur va tures. The trans verse co or di nate q3 is reck oned to ward the in creas ing ex ter nal nor mal to the sur face W (Fig. 1). Let e1 and e2 be the unit vec tors of tan gents to the co or di nate lines q1 and q2 , a e3 3= the unit vec tor of the ex ter nal nor mal, Aa Lam\u00e9 pa ram e ters, ka the main cur va tures, d A the dis tances from the ref er ence to ex ter nal sur faces W A , r( , )q q1 2 the ra dius vec tor of the ref er ence sur face W, RM ( , )q q1 2 the ra dius vec tor of the midsurface W M , and RA ( , )q q1 2 the ra dii vectores of the ex ter nal sur faces W A . Here in af ter, = 1, 2,..., NL; a, b = 1, 2; i, j, l, m = 1, 2, 3; A = - +, . For the ra dii vec tores of the mid dle and ex ter nal sur faces of the shell, we have R r aI Iz= + 3 ( , , )I = - +M , (1) z d- -= - , z d+ += , z z zM = +- +1 2 ( ). wherefrom fol low the for mu las for base vec tors (see Fig. 1) a r ea a a a= =, A , g R ea a a a a I I IA c= =, , (2) c k zI I a a= +1 ( , , )I = - +M . Let us as sume that the tan gen tial and trans verse dis place ments are dis trib uted across the shell thick ness ac cord ing to the lin ear and square laws, re spec tively [15], i.e., u N ua a= \u00e5 A A A , u L uI I I 3 3 = \u00e5 ( , , )I = - +M , (3) where ua q qA ( , )1 2 and u 3 1 2 A ( , )q q are the tan gen tial and trans verse dis place ments of the ex ter nal sur faces; u 3 1 2 M ( , )q q is the trans verse dis place ment of the midsurface; N A ( )q3 and LI ( )q3 are the Lagrange poly no mi als of the first and sec ond de grees, re spec tively: N h z- += - 1 3( )q , N h z+ -= - 1 3( )q , (4) L N N N- - - += -( ), L N NM = - +4 , L N N N+ = -+ + -( )"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002756_wst.2010.832-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002756_wst.2010.832-Figure2-1.png",
+        "caption": "Figure 2 | Schematic diagram of the pilot filter.",
+        "texts": [
+            " The use of co substrates can substantially increase the transformation rate of the dye to amines under anaerobic conditions and consequently the mineralization of the resulting amines under aerobic conditions (Field et al. 1995; Razo-Flores et al. 1997) (Figure 1). The main objective of this research was to determine the feasibility to treat municipal wastewater containing dissolved azo dye direct blue 2 in a biological aerated filter (BAF) using lava stones as support of the microorganisms and under combined anaerobic/aerobic conditions. A pilot filter was built according to Figure 2 and Table 1. The filter is a PVC pipe, with an internal diameter of 19 cm, in vertical position filled with a 3.0 m lava-stone bed. The lava stones were sieved to a size of 6 mm. Table 2 shows the main characteristics of the filter media. The experimental procedure was divided in two parts: During the first part of the experiment, under continuous aeration, the filter was fed with municipal wastewater from a neighboring settlement combined with the wastewater from the main campus of the National University of Mexico and, during the second part of the experiment, the azo dye direct blue 2 was dosed to the wastewater to yield a final concentration in the wastewater of 50 mg/l (34 mgCOD/L) (O\u2019Neill et al"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002877_j.phpro.2012.10.047-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002877_j.phpro.2012.10.047-Figure3-1.png",
+        "caption": "Fig. 3. Laser processing unit",
+        "texts": [
+            " The scanner-based process combines the advantages of conventional laser cladding, as shown in figure 2, with a highly enhanced efficiency by a significant increase of the melting deposition rate, improved process stability, quality and a flexible weld seam geometry [8]. For scanner-based cladding and hardening, a modular processing unit was built and integrated into a hybrid machining center prototype. The same unit can also be used for laser hardening, so that for both processes only a single optical setup and laser is required. In this way, investment costs are reduced and a flexible change between hardening and cladding without the need for reclamping is possible. The setup of the processing unit is illustrated in figure 3. The laser beam is guided over the workpiece by a 2-D laser scanner. A telecentric f-theta lens allows for the processing of a planar scanning field while maintaining orthogonal beam incidence. The motion of the laser spot within the scanning field can be flexibly controlled by software. This allows for the variable adjustment of the seam or track width for both applications, hardening and cladding. The wire feeding system for the supply of additive material has a hot-wire functionality which may be used to stabilize the melting process by local wire preheating",
+            " The ability of programming a larger set of process parameters introduces the requirement of specifying appropriate parameter combinations in every program. The programmed scanner motion must be chosen accordingly with the laser power, filler wire rate and process speed in order to succeed in reliably producing a welding bead geometry that was found during process qualification. With process data like the operation point shown in figure 4, the amount of process data that needs to be programmed increases significantly. By integrating the laser processing unit (figure 3) and an industrial robot into a multi-technology platform [9], a CNC (Computerized Numerical Control) system was available to serve as a master control system. With the laser processing unit and the robot being linked to the master control as slaves, process control functions could be implemented on the CNC. It could also provide programmer support functionality at a single point of access. Each set of parameters that enables a specific laser cladding process was stored in a data structure which effectively encapsulates an entire laser cladding operation point"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001359_electromotion.2009.5259143-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001359_electromotion.2009.5259143-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of the proposed method",
+        "texts": [
+            " Negrea has also studied this subject in [2]. G.A. Capolino, J.F. Brudny and D. Belkhayat have also presented valuable studies in measuring the leakage flux by an external sensor [3]-[5]. All these studies aim to improve the condition monitoring for electrical machines. They try to capture harmonics caused by machine faults such as broken rotor bars, worn bearings etc. by axial or radial leakage flux sensors. But none of them try to estimate the rotor speed, position and number of rotor slots by an external sensor. Let consider Fig. 1, which is a sketch of the magnetic circuit of a typical motor. It is clear that some of the mutual flux is going to leak out of the magnetic circuit. Note that this flux is revolving at the synchronous speed. Therefore, it may be argued that if a search coil is placed outside the frame; an induced voltage containing the information available in the mutual flux will be obtained. 978-1-4244-5152-4/09/$26.00 \u00a92009 IEEE It may be possible to process this signal to obtain the rotor speed and rotor slot number or any other information available in the mutual flux"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000117_iros.2006.282358-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000117_iros.2006.282358-Figure1-1.png",
+        "caption": "Figure 1: All-Wheel Steering Mobility System. For mobile robots with the property that all of the wheels are steerable (e.g. Rocky 8, shown), the body can rotate about any point in the local tangent plane.",
+        "texts": [
+            " Ackerman and differential drive methods require that the velocity vector in the local tangent plane be aligned with the forward axis of the robot and the center of rotation about along an axis aligned with the fixed set of wheels. Omnidirectional methods allow independent motion in translation and rotation [8]. Vehicles with all-wheel steering capability differ from these models in that the velocity vector can be aimed in any direction in the local tangent plane and the center of rotation can be anywhere in the local tangent plane (Figure 1), but the path heading cannot change arbitrarily due to the nonholnomic constraints of the wheels. 1.2 Related Work In the context of robot motion planning, most research in trajectory generation has dealt with finding obstacle-free paths subject to nonholnomic constraints assuming flat terrain and simple vehicle models. Most of the work to date falls into one of two categories: graph search via a sequence of low-order geometric primitives or optimization using a single high-order primitive. Some of the first work in trajectory generation involved composing optimal paths from sequences of line segment arcs [9], clothoids [2][11], and cubic spirals [3]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003812_robio.2012.6491108-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003812_robio.2012.6491108-Figure7-1.png",
+        "caption": "Fig. 7. Angles and hysteresis for ankle and ground contact for GA changes.",
+        "texts": [
+            " Any landing position above 200 mm or below -200 mm should be understood as the limb climbing the slopes. After comparative data between IL and ST, the influence of gastrocnemius on directional change was measured. Setting IL and ST pressures at 0.25 MPa and 0.15 MPa, respectively, we calculated the average and standard deviation of 25 jumps for 6 different pressures at GA, as shown in Fig. 6. Following this experiment, we measured the contribution of GA pressure to angle variation at ankle and ground contact, as seen in Fig. 7. The left dashed line at the graph indicate the point where the transition from digitigrade to plantigrade stance happens, while the right dashed line indicates angles where the hindlimb collapses. The criteria for defining these two thresholds is when the ankle touches the floor and when the angle at the ankle reaches 160 degrees, respectively. Using data from previous figures, we estimated a relationship between ST pressure and landing position. Owing to better results compared to the polynomial approach, we decided to use a Boltzmann based correlation equation of the available data set: X(pST ) = A2 + A1 \u2212A2 1+ e(pST\u2212pST 0)/d pST (7) where XpST is the landing position, A1 the lower limit, A2 the upper limit, pST 0 is approximately half of the pressure amplitude and d pST is the pressure range",
+            " On the GA muscle, the opposite phenomenon was observed. Increasing GA pressure resulted in a negative foot placement, and we suggest that the extension of the ankle combined with the flexion of the knee brings the center of gravity of the leg to the back, as seen in Fig. 6, thus rotating the system backwards. Although it would be possible to control the hopping direction through gastrocnemius pressure, altering gastrocnemius pressure has a critical shortcoming for the hindlimb stability: As seen in Fig. 7, different pressures for gastrocnemius result in different attack angles. At one extreme, we have a plantigrade landing, which is not very efficient dissipating potential energy from the jump, while at another we have an ankle joint locked at 180 degrees, which simplifies the leg as a 2-link system, being also degrading for energy dissipation. Although biological data on how cats change muscular activation for different jumping directions is scarce, we propose that different tension on the ST muscle largely contributes to a better control over feline leaps"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000544_1.2991175-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000544_1.2991175-Figure3-1.png",
+        "caption": "Fig. 3 Lemon \u201eleft\u2026 and barrel \u201eright\u2026 shapes",
+        "texts": [
+            " 1 In full active zone, D = p \u2212 pcav, D 0 B = 1 2 In cavitated non-active zone, D = r \u2212 h, D 0 B = 0 ithout misalignment, the film thickness is given by h ,y,t = h0 ,t + he ,y,t + ht ,y,t + hd ,y,t 4 here =x /R is the angular coordinate for a housing of R radius nd h0 , t is the nominal film thickness; for a circular bearing 0 , t =c 1\u2212 x t cos \u2212 y t sin , he ,y , t is the elastic deormation of the bearing housing and shaft due to the hydrodyamic pressure, ht ,y , t is the deformation due to thermal exansion of the bearing housing and shaft, and hd ,y , t defines a arrel and/or a lemonlike shape, explained on Fig. 3. The lubricant viscosity is assumed to vary with the temperature T = 0e\u2212 T\u2212T0 + a 5 here a is an asymptotic viscosity, 0 is the oil viscosity at T0, nd is the thermoviscosity coefficient. It is also supposed to be ressure dependent according to the Barus equation, and then P,T = T e P\u2212P0 6 here is the piezoviscosity coefficient and P0 is the ambient ressure. To determine the mean temperature T, the power resulting from uid shearing Pf is considered to be 80% evacuated by the fluid tself value commonly admitted for engine bearings and 20% hrough the solids",
+            "3 Connecting rod structural density kg /m3 7800 Connecting rod length, center-to-center mm 144 Mass kg 0.486 Center of mass mm 35.35 Polar moment of inertia kg m2 2.32 1003 Table 4 Other bearing specifications Diameter mm 48 Surface roughness m 0 Liquid density kg /m3 900 Ambient pressure MPa 0 gauge Cavitation pressure MPa \u22120.1 gauge Table 5 Lubricant properties Level \u2212 Level + Transactions of the ASME of Use: http://www.asme.org/about-asme/terms-of-use i f = s j n d t s i r c t t 6 T d 1 o s t t d g t l a w i g o c e e S R L B J Downloaded Fr The lemonlike shape, Fig. 3, is introduced to take the bolt load nto account and other defects of the same kind. The range starts rom a perfect circle, c =0, to a 10 m deformed circle, c 10 m. The empirical development of engine connecting rods has reulted in relief on the back of the bearing to deflect away from the ournal thereby counteracting the tendency for the oil film thickess to diminish toward the edges, Fig. 3. Indeed, to prevent some amages on the shell border, due to contact occurrences between he shaft and the housing, a barrel shape could be realized on the haft. However, for manufacture reasons, it is more convenient to ntroduce this correction on the shells; the film thickness field emains the same. The maximum value retained is c =2 m beause higher values cause pressure spikes and a very thin film on he bearing central zone. These spikes, conjugated with low film hickness, can induce local damage on the shells, as shown on Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000220_1.2424240-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000220_1.2424240-Figure5-1.png",
+        "caption": "Fig. 5 Principal membrane force field \u201ePKII, plotted onto the deformed configuration of a The figure highlights the region where th \u22121.49E\u201303 kPa cm\u2026.",
+        "texts": [
+            "org/ on 05/20/201 Figure 4 c shows directions of the maximum principal Cauchy membrane forces that in the wrinkled regions indicate the directions of the wrinkles horizontal in this sheet . In this problem, wrinkling is generated in the sheet by the action of the clamped boundaries that prevent transverse contraction of the sheet and set up a local biaxial state of stress; the transverse stresses are tensile near the boundary but they would tend to be compressive away from it. These stresses are shown in the diagram of Fig. 5 that has been obtained using the constitutive model Eq. 2 that allows compressive forces to develop in the sheet. As for the previous example, regions of behavior similar to those of Fig. 4 have been observed in Fig. 1 b in the thin polyethylene membrane tested in Ref. 2 and subject to similar boundary and loading conditions. 6.3 Simulation of Reconstructive Surgery. In 7,28 a numerical procedure has been formulated for the simulation of reconstructive surgery of the skin. The procedure refers to surgical operations characterized by the following steps: the incision of the skin; the undermining of a portion of the skin surrounding the cutaneous defect cancerous and noncancerous growths, burn wounds, lacerations, birth defects ; the excision of the defect; the closure and suture of the wound"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001399_1528083707087832-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001399_1528083707087832-Figure2-1.png",
+        "caption": "FIGURE 2. Tensile test \u2013 tubed method.",
+        "texts": [
+            " DIN EN 29073-T3 (German version of the international standard ISO 9073-3) using strip samples (untubed samples) for tensile test. Five specimens were tested at angles of 0 (machine direction) and 90 (cross-machine direction). Listed in Table 1 are some testing parameters for a brief comparison between the two tensile tests for the nonwovens. Based on the previous research [21], the method of DIN 53857-T2 usually produces much more reliable results. So, the use of the above both methods may provide a possibility to compare measured data from the different testing methods. Figure 2 illustrates the mounting of a tubed nonwoven specimen. Thickness of the nonwovens was measured at GEORGE WASHINGTON UNIVERSITY on January 24, 2015jit.sagepub.comDownloaded from according to DIN 53855. Mass of the samples was tested using the standard DIN EN 12127. To allow a direct comparison among the tested samples with different material density (or mass), the aforementioned tensile tests use a normalized tensile stress to describe the material strength. This strength is defined as: \u00bc P \u00bc F b t \u00bc F bw N tex \u00f01\u00de where, P\u00bcmaximum tensile stress (N/m2); F\u00bcmaximum force (N); \u00bc tested sample density (g/m3); t\u00bc tested sample thickness (m); at GEORGE WASHINGTON UNIVERSITY on January 24, 2015jit"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001407_iccet.2009.41-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001407_iccet.2009.41-Figure1-1.png",
+        "caption": "Fig 1. Model of FEM mesh",
+        "texts": [
+            " In accordance with the above-mentioned requirements of the motorized spindle model, dualdimensional plastic beam element \u201cBEAM188\u201d is chosen to simulate the whole shaft, and spring COMBIN14(damp element) is chosen to simulate the bearings. In accordance with the static design requirements, the rated rotating speed and the maximum rotating speed of the motorized spindle are designed as 12000 r/min and 18000 r/min respectively in the research. 2.3.2. Loading and Calculating. Zero-bound is the only load in the model analysis. The front bearing (bearing one) of the motorized spindle designed in this paper is a fixed-end which constrains the degrees of freedom of Ux, Uy and Uz shown in figure 1. The rear bearing, that is bearing four(as shown in figure.1), is constrained the same way as the front bearing in the direction of Ux although there is slightly displacement along the direction of the axis. The axis with the constrain and grid is shown in figure 1. 2.3.3. Result. Table 1 shows the first seven order natural frequency of the axis of the motorized spindle after the calculation of ANSYS: Harmonic analysis method is utilized to analyze the periodic response to the continuous periodic load on the structure system, determine the steady-state response of the linear structure when the linear structure bears the load following the regularity of Sine. The forced vibration equation of n-degrees of freedom without damping system is expressed as follows: [ ] ( ){ } [ ] ( ){ } [ ] ( ){ } ( ){ }M x t C x t K x t P t+ + = \uff084\uff09 Where, ( ){ } ( ) ( ) ( ) ( ){ }1 2 3 nP t p t p t p t p t= is any exciting force"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001362_tro.2008.2008745-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001362_tro.2008.2008745-Figure1-1.png",
+        "caption": "Fig. 1. Gough\u2013Stewart platform observed by a classical perspective camera. (a) Camera position with respect to the platform and (b) image of the legs. A Gough\u2013Stewart platform observed by an omnidirectional camera. (c) Camera position with respect to the platform and (d) image of the legs.",
+        "texts": [
+            "00 \u00a9 2009 IEEE direct application of visual servoing techniques assumes implicitly that the robot inverse differential kinematic model is given and that it is calibrated. Therefore, [1] and [2] propose, respectively, image-based and position-based visual-servo schemes by directly observing the platform legs with a classical perspective camera. Unfortunately, to position adequately the camera to observe simultaneously all the platform legs is a complex task. The camera was positioned in [1] and [2] in front of the platform [see Fig. 1(a)]. In this case, the legs in front of the platform are closer to the camera than the ones in the back. As a consequence, the extraction of the image features lying on legs in the back will be less robust. Furthermore, large parts of the legs in the back are occluded by the front legs [see Fig. 1(b)] and full occlusions can happen. This is an important drawback since the vision-based control assumes that all legs can be observed during the servoing task. A first solution to address this issue could be to employ a system made of multiple cameras. However, in this case, data provided by each camera must be synchronized and the multicamera system calibrated. A second and simpler solution, whose first results were presented in [14], consists of positioning a single omnidirectional camera (vision system providing 360\u25e6 panoramic views of the scene) at the platform center [see Fig. 1(c)]. This way, all the legs can be simultaneously observed in a panoramic view, and potential occlusions cannot occur [see Fig. 1(d)]. Moreover, by positionning the omnidirectional camera at the platform center, the feature extraction should be more robust than when a conventional camera is employed since the legs will be closer to the image plane. Finally, observing legs, even using an omnidirectional camera allows a linear calibration of the platform [6]. Clearly, visual servoing of the Gough\u2013Stewart platform will benefit from the enhanced field of view provided by an omnidirectional camera. However, omnidirectional images exhibit supplementary difficulties compared with conventional perspective image (for example, the projection of a line is no more a line but a conic curve)",
+            " The world point X is projected in the image plane into the point of homogeneous coordinates p = Km, where K is a 3 \u00d7 3 upper triangular matrix containing the conventional camera-intrinsic parameters coupled with mirror-intrinsic parameters and m = [x y 1] = [ X Z + \u03be\u2016X\u2016 Y Z + \u03be\u2016X\u2016 1 ] (1) The matrix K and the parameter \u03be can be obtained after calibration using, for example, the methods proposed in [13]. In the sequel, the central imaging system is considered calibrated. In this case, the inverse projection onto the unit sphere Xm can be obtained as Xm = \u03bb [ x y 1 \u2212 \u03be \u03bb ] (2) where \u03bb = \u03be+ \u221a 1+ (1\u2212\u03be 2 )(x 2 + y 2 ) x 2 + y 2 +1 . A Gough\u2013Stewart platform has six cylindrical legs of varying length qj (j = 1, . . . , 6) attached to the base by spherical joints located at points Aj , and to the moving platform by spherical joints located at points Bj (see Fig. 1). The image of the jth leg is defined by the projection onto the image plane of two lines (Lj 1 and Lj 2 ), as depicted in Fig. 2. Let ni j = [ni jx ni j x ni j x ] (i = 1, 2) be the unitary vector orthogonal to the interpretation plane \u03c0i j defined by the line Li j and the projection center. The points Xm lying on the intersection between \u03c0i j and the sphere are then defined by{ \u2016Xm \u2016 = 1 ni j Xm = 0. (3) Using the spherical coordinates given by (2), it can be shown that 3-D points lying on Li j are mapped onto points m lying on a conic curve \u0393i j , which can be written as \u03b10x 2 + \u03b11y 2 + 2\u03b12xy + 2\u03b13x + 2\u03b14y + \u03b15 = 0 (4) where \u03b10 = ni2 j x \u2212 \u03be2 (1 \u2212 ni2 j y ), \u03b11 = ni2 j y \u2212 \u03be2 (1 \u2212 ni2 j x ), \u03b12 = ni jxni j y (1 \u2212 \u03be2 ), \u03b13 = ni jxni j z , \u03b14 = ni j y ni j z , and \u03b15 = ni2 j z "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001451_978-0-387-09643-8-Figure9.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001451_978-0-387-09643-8-Figure9.1-1.png",
+        "caption": "Figure 9.1: Visualization of a Snow Monitoring Scenario. (a) The individual sensor nodes collect local information about the snow state, e.g., temperature, light intensity, pressure, or humidity. (b) The observations provide useful information for snow avalanche warning systems and allow an effective utilization of water resources.(adapted from [132])",
+        "texts": [
+            " Each year, snow avalanches cause casualties and damage, not only in non-protected areas but also in popular cross-country skiing areas, e.g., in the Wasatch mountains in Utah. The application of an intelligent and autonomous sensor network could offer useful information for the support of avalanche forecasting systems. The individual sensor nodes deployed on the ground or within the snow pack collect measurable information about the snow state, such as temperature, light intensity, pressure or humidity; see Figure 9.1 (a). Then, based on these observations and after further processing, measures about the stability of the snow pack, e.g., stress distribution, strain distribution, density distribution or location of so-called weak layers, of a certain area could be estimated [9, 99, 100, 127, 151]. Thus, by means of a sensor network, the possibility of snow avalanches can be predicted and defense structures in avalanche starting zones can be optimized; see Figure 9.1 (b). An additional application scenario where sensor networks could provide novel possibilities is the accurate and efficient evaluation of snowmelt. By this means, water resources could be utilized more efficiently and flood runoffs could be forecast more accurately [76, 95]. A further example worth mentioning is the application of sensor networks to monitoring the condition and composition of ice in skating rinks [171]. For speed skaters to reach faster times, the optimal ice composition and especially the optimal temperature distribution of the ice is quite essential",
+            " Each year, snow avalanches cause casualties and damage, not only in non-protected areas but also in popular cross-country skiing areas, e.g., in the Wasatch mountains in Utah. The application of an intelligent and autonomous sensor network could offer useful information for the support of avalanche forecasting systems. The individual sensor nodes deployed on the ground or within the snow pack collect measurable information about the snow state, such as temperature, light intensity, pressure or humidity; see Figure 9.1 (a). Then, based on these observations and after further processing, measures about the stability of the snow pack, e.g., stress distribution, strain distribution, density distribution or location of so-called weak layers, of a certain area could be estimated [9, 99, 100, 127, 151]. Thus, by means of a sensor network, the possibility of snow avalanches can be predicted and defense structures in avalanche starting zones can be optimized; see Figure 9.1 (b). An additional application scenario where sensor networks could provide novel possibilities is the accurate and efficient evaluation of snowmelt. By this means, water resources could be utilized more efficiently and flood runoffs could be forecast more accurately [76, 95]. A further example worth mentioning is the application of sensor networks to monitoring the condition and composition of ice in skating rinks [171]. For speed skaters to reach faster times, the optimal ice composition and especially the optimal temperature distribution of the ice is quite essential"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001359_electromotion.2009.5259143-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001359_electromotion.2009.5259143-Figure2-1.png",
+        "caption": "Fig. 2. Three different search coils tested in this research. The coil placement and the flux path are shown",
+        "texts": [
+            " Therefore, it may be argued that if a search coil is placed outside the frame; an induced voltage containing the information available in the mutual flux will be obtained. 978-1-4244-5152-4/09/$26.00 \u00a92009 IEEE It may be possible to process this signal to obtain the rotor speed and rotor slot number or any other information available in the mutual flux. It is clear that the first stage of such a research is making sure that the shape of the search coil core is such that the search coil voltage due to residual flux is maximized. Three types of search coil cores are experimented with in this work. Fig. 2 displays how these coils are placed on the frame of several induction motors with cast iron frame. In Fig. 3, photograph of the U-shaped coil placed on the frame of an induction motor is shown. Induced voltage in different coils is measured in various axial and radial positions. It is found that the radial position of the coil is not critical. However, placing correctly the coil in the axial direction is very important. In Fig.3, the effect of the axial position of the search coil on the induced voltage is displayed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000863_1-84628-484-8_2-Figure2.16-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000863_1-84628-484-8_2-Figure2.16-1.png",
+        "caption": "Figure 2.16 A schematic of the MORE device.",
+        "texts": [
+            " Similar strategies can be adopted for integrated sensors where the reference electrode can be constructed by chloridising a silver plated pin or pad close to the working electrode surface. Additional information about circuits and sensor ancillaries, complete with recipes, has been published by Cass [45]. A new type of electrochemical device, the Micro-Optical Ring Electrode, or MORE was introduced in 1996 [46]. It consists of a fibre optic light guide on which a thin film of gold has been deposited as shown in Figure 2.16. The gold is anchored by chemical bonding to thiol terminated silyl groups to obviate the need for a chromium underlayer. The idea is to provide intimate relationship between a light source and the electrode surface which can, unlike optically transparent electrodes, be readily renewed by conventional lapping processes. There are some redox reactions which take place more rapidly if the reactant is in a photoexcited state. If the electrode potential is poised between the energy of Highest Occupied Molecular Orbital (HOMO) and the Lowest Unoccupied Molecular Orbital (LUMO), the molecule can undergo neither oxidation nor reduction since the HOMO is full and the LUMO is of a higher energy"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002967_amr.317-319.2226-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002967_amr.317-319.2226-Figure1-1.png",
+        "caption": "Figure 1 Gear reducers for electric elevator",
+        "texts": [],
+        "surrounding_texts": [
+            "Generally, the electric power system of machinery consists of motor and gear reducer to get large output torque. The gear reducer of the power system for elevator can be ordinary spur gear reducer, worm and worm gear reducer, planetary gear reducer, or coupled type gear reducer. Planetary gear trains are commonly used in various transmissions due to the reason of compact size, light weight, and multi-degrees of freedom. One of its applications is used as power systems of elevator. Planetary gear trains were the subject of intensive research directed at kinematic analysis [1-2], kinematic design [3-5], and efficiency analysis [6-8]. Only some studies focused on the kinematic design and efficiency analysis of planetary gear trains for specific purpose. This paper focused on the kinematic design and efficiency analysis planetary gear reducer for elevator."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure31-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure31-1.png",
+        "caption": "Fig. 31. Reduced stresses in sheets during riveting (pressing stage) a) w1 b) w2",
+        "texts": [
+            " The plastic deformation region of sheets after upsetting for two cases of loading is shown in Fig. 29. Decreasing of height of the formed rivet head during riveting process causes the increasing of plastically deformed area around the rivet hole (Fig. 29a, b). Methods for FEM analysis of riveted joints of thin walled aircraft structures 963 Negative stresses (compressive stress state), considerably exceeding yield stress level, appear in the material of sheets around the rivet hole (Fig. 30). Reduced stress distribution during riveting (pressing) is presented in Fig. 31. Jerzi Kaniowski et al. 964 Riveting process has a significant impact on the stress state in joined metal sheets, especially in the case of uniform filling of the rivet hole by the rivet shank. A certain method of material transfer from rivet head to rivet shank is the application of rivets with compensator [OST 1 34040-79]. Uniform distribution of residual stresses in the rivet hole increases fatigue life of the rivet joint. Residual stress values should be adjusted to the average level of operating stresses in cover panels of the considered aircraft"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002171_iccis.2010.265-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002171_iccis.2010.265-Figure1-1.png",
+        "caption": "Figure 1. Structure diagram of the ball and beam system",
+        "texts": [
+            " during the input and output, a system model is constructed with the application of the control method concerning adaptive neural network, and then, an analysis and research on the control performance of the system is given, which consequently leads a fairly good control effect, and also is meaningful for the groping for control methods on complex systems and the solving of issues concerning robustness and stability of the system in the control process. II. MODEL BUILDING OF THE BALL AND BEAM SYSTEM As shown in Fig.1, the ball and beam system can be divided into two large parts which are executive system and control system. The executive system is a four-bar linkage with the bar composed by a steel scale bar and a linear displacement transducer; the bar can rotate about its left pivot and the control system is motivated by a DC servo motor. Through control on the angle and swing speed of the bar, the balance of globule on a certain set point of the bar is achieved. In accordance with the principle of dynamics and Lagrange equations, it can be concluded as below [2]: if, in a dynamic system, the kinetic energy is T , potential energy is V , and the energy dissipation function is R , then, the motion law of the system can be represented with Lagrange differential equation as follows: 978-0-7695-4270-6/10 $26"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001838_cca.2010.5611256-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001838_cca.2010.5611256-Figure2-1.png",
+        "caption": "Fig. 2. Coordinate of the quadrotor",
+        "texts": [
+            " The quadrotor to be considered, which is shown in Fig. 1, has 4 propellers, 3 of which are horizontally mounted to control its pitch and roll rotations while the last one is vertically mounted to control its yaw rotation. This quadrotor has 3 outputs and 4 inputs. The outputs are the pitch angle, roll angle, and yaw angle, while the inputs are 978-1-4244-5363-4/10/$26.00 \u00a92010 IEEE 1731 the control voltages of the 4 propellers\u2019 motors equipped at the 4 ends of the quadrotor. The coordinate of the quadrotor is shown in Fig. 2, where Fx(x = f, l, r, b) denotes the thrust forces generated by 4 propellers, its suffixes mean its locations which are front, left, right, and back. The structure is assumed to be symmetrical, the origin is assumed to coincide with the quadrotor\u2019s centroid. According to the kinetic equation of each axis, 3 differential equations can be formulated under the condition that the quadrotor is horizontally postured [6]. The pitch p is defined to be the angle circled around the Y axis, and the counterclockwise rotation round Y axis is defined to be positive"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003727_j.engfailanal.2011.06.004-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003727_j.engfailanal.2011.06.004-Figure7-1.png",
+        "caption": "Fig. 7. (a) Order of the U-bolts placement in fan assembly. (b\u2013e) Equivalent von-Mises stress distribution on the U-bolts for four different orientations of fan blade respect to the direction of gravity.",
+        "texts": [
+            " The transition from these stages is revealed on SEM micrograph of Fig. 6d. The fracture mode at final stage may be either ductile (with a dimpled fracture surface) or brittle (with cleavage or intergranular fracture surface) or a combination of both [12]. Fig. 6e shows SEM image of the fracture surface at the final stage. The presence of dimples indicates that the fracture mode at this stage is ductile. Stress analysis is done for 16 directions of blade in vertical plane by finite element analysis. The order of U-bolts placement in fan assembly is shown in Fig. 7a. Fig. 7b\u2013e illustrates equivalent von Mises nodal stress distribution for particular four situations of blade having angles of 0 , 90 , 180 and 270 respect to gravity direction. It is observed that the nodes have the high stress values near the nuts and below the contact zone with hub plate. The maximum and minimum stress states are detected in Fig. 7c and e, respectively; that is in situations 90 and 270 . It is noted that the maximum nodal stress is calculated in the second bolt of the fan assembly shown in Fig. 7a. Because of the complexity of loading condition, a multiaxial state of stress is formed in U-bolts. Since, the greatest Mises nodal stress is less than the material yield strength, (Sy = 205 MPa), general yielding cannot occur in this analysis. This is consistent with the results of fractography examinations obtained in the previous sections. Since the bolt is actually loaded with dynamic forces, fatigue analysis is performed. Because the applied stresses are within the elastic range of the material, the stress-life approach is used to determine the fatigue life of U-bolts [11]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure3-1.png",
+        "caption": "Fig. 3. PZL M28 wing schema with marked analyzed region [11]",
+        "texts": [
+            " The airplane wing consists of a rectangle centre wing and two trapezium outer wing parts. It features the 3\u00ba anhedral and 4\u00ba geometric twist. The aim of the work was to analyze low skin riveted joint near rib no 21 on the PZL M28 Skytruck wing. In the case of thin wall consisting of frame (ribs, stringers) and skin, it is assumed (for finite rib stiffness) that force system closes in two bulkheads (between three ribs), on each side. In order to put loads from rest of wing correctly, the model includes three bulkheads on each side of rib no 21, i.e. part between ribs no 18 and 24 (fig. 3). Thanks to that, there is the correct stress distribution on bulkheads near rib no 21. Methods for FEM analysis of riveted joints of thin walled aircraft structures 945 The wing is a torsion box construction. The wing torsion is made from spars, ribs and skin panels stiffened with stringers. The spars consist of upper and lower flanges of T-section and webs reinforced with struts. Ribs consist of upper and lower flanges, webs reinforced with struts and connectors. The part of wing FEM model is shown in figure 4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000689_09544100jaero206-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000689_09544100jaero206-Figure1-1.png",
+        "caption": "Fig. 1 Schematic configuration of CTAT",
+        "texts": [
+            " Trim algorithm is developed in order to learn the range of the required thrust-vector angle to attain a feasible steady state flight in different flight phases. The obtained model is then used towards an understanding and investigation of the dynamic stability and consequently, flying qualities of the aircraft. In addition, transition manoeuvre in take-off is numerically studied and the vehicle\u2019s stability is described by examining the system poles over the established trajectory. Finally, Matlab/Simulink simulation environment [26, 27] is used to review the trim algorithm and dynamic behaviour. The CTAT schematic configuration is illustrated in Fig. 1. Thanks to the conventional design of Banshee, physical data, and mass properties are wellestimated with the available information [28]. The conventional control surfaces are assumed to be fixed and the piston-prop engine is virtually substituted by two small gas turbines [29]. Here, for the sake of simplicity and completeness, only the major finalized characteristics of the aircraft are listed in Table 1. It is assumed that the thrust deflection angle in pitch direction (\u03b1T) could be different for the left and right engines but in yaw direction (\u03b2T) has to be identical for two engines and the positive directions are in accordance with the right-hand rule, respecting the body-axis (positive \u03b1T causes the nose pitching down and positive \u03b2T causes the nose turn left)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002418_10402004.2010.491174-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002418_10402004.2010.491174-Figure3-1.png",
+        "caption": "Fig. 3\u2014VSFT test steel annulus and friction material plate configuration.",
+        "texts": [
+            " When running ATF tests, the lubrication regime is in the boundary or mixed regime (at high speed, typically starts from 350\u2013500 rpm). The force is applied to the test components by a double-acting pneumatic cylinder and measured on a load cell with 445 N capacity. A digital servomotor drives the upper test component in the range of 1 to 2,400 rpm. Motor torque capacity is 10.7 Nm at 8.3 rpm and 8.8 Nm at 2,400 rpm. Torque is measured by a rotary transformer within the range of 0 to 22.6 Nm, with a resolution better than 0.113 Nm. Figure 3 shows a schematic representation of the test piece configuration, and Fig. 4 and Table 1 show the structure of the test procedures. The test started from one set of the LVFA speed sweep tests at three controlled temperatures. Then the sample was subjected to a 2-h durability phase under constant speed, load, and temperature. Then the LVFA test cycle was repeated. This sequence was repeated for a total of nine LVFA speed sweep runs. The pressure applied was 2.12 MPa. The annulus was on the rotor component, and the disk that carried the friction mate- Start LVFA speed sweeps @ 3 temperatures (27, 66 and 121\u00baC) 2 hours durability (ageing) test at constant speed, load & temperature End Repeat until 16 hours total durability Fig. 4\u2014Structure of VSFT test. rial was flat, as schematically shown in Fig. 3. Components were immersed in 50 mL of test oil. Friction force was measured continuously during the test. As mentioned earlier, four model ATF formulations were tested in this study. They all used the same fully formulated dispersant/inhibitor package, but different friction modifier/detergent systems were used for each of the four formulations to investigate the effects on the friction behavior. The details of the formulations are shown in Table 2. II.4 oil was formulated without calcium sulfonate detergent"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001872_1.4001203-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001872_1.4001203-Figure1-1.png",
+        "caption": "Fig. 1 Nonsingular and singular postures of a mobile planar four-bar linkage, and a singular posture of an immobile planar",
+        "texts": [
+            " ournal of Mechanical Design om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/27/201 When this occurs, the quadratic equation in Eq. 36 has roots for t or alternatively t\u22121, giving the solutions for vector . 6.1 Four-Bar Linkage. A planar four-revolute, four-bar linkage may be treated as a single-loop platform mechanism, where the coupler is the platform, connected to the ground on each side by R-R dyads. All four revolutes share the same rotation axis direction along the z-axis. Figure 1 depicts a mobile parallelogram linkage, in nonsingular and collapsed singular postures, along with one posture of an immobile linkage, with numeric results presented in Tables 1\u20133. The first two rows in Table 1 give the twists for the platform connection to the ground through the crank link, while the next two rows give the twists for the connection to the ground through the follower link. The mobile linkage in the nonsingular posture has the well-known single degree-of-freedom for a four-bar linkage, a pure translation at right angles to the common orientation of the crank, and follower links in the case of a parallelogram linkage"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure7-1.png",
+        "caption": "Figure 7. S-path when \u03b8=0.",
+        "texts": [
+            " (6) Case 2-2 (Py<0 and \u03b8=0) A special case arises when the target direction is parallel to the current direction, i.e., \u03b8=0. First, the mid point of the line segment that connects the current position and target position is denoted as M, and its coordinate position as M(Mx,My). Then, the goal of two circles of maximum radius is to find, first, a circle which passes through the mid point M and is tangent to the current direction line, and second, a circle which passes the mid-point M and is tangent to the target direction line, as shown in Figure 7. Only a small amount of geometric manipulation is required to derive the equation for R4 as follows; (7) The case of \u03b8<0 when the target position is in the first quadrant is reserved for the reader\u2019s preference. Case 3: Employment of Minimum Radius Path If the radius of an arc path obtained in either Case 1 or Case 2 is smaller than the Minimum Radius, denoted by Rmin the minimum radius a vehicle can follow, (normally about 4.5 meters), an additional arc should be added to the Rmin arc path. Figure 8 presents a case where the target position is quite close to the Yw axis and the directional difference \u03b8 value is large"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002546_s1052618810050018-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002546_s1052618810050018-Figure3-1.png",
+        "caption": "Fig. 3. Mechanism with four degrees of freedom and cardan shaft: (a) scheme of the mechanism; (b) system of kinematic screws in the usual configuration; (c) system of kinematic screws in the special provision.",
+        "texts": [
+            " Moreover, as in the previous case, in the third kinematic chain, the first link of the first parallelogram and the final link of the second are linked, respectively, with the rotational drive and the output link mechanism in the cen ters of these links, and the final link of the first parallelogram coincides with the initial link of the second. Here, the rotary engine provides the orientation of the output link. The gear ratio of the rotation is unity. Consider a mechanism in which two kinematic chains are imposed on the two connections (Fig. 3), and the third does not impose the connections. The first and second kinematic chains, as in the previous case, consist of one drive rotational pair (rotary engine) located on the base, one intermediate rotational pair located with the axis parallel to the axis of rotation of the drive, and a finite cylindrical two mobile pair (the axes of the finite cylinder pairs of the two chains are the same). The third kinematic chain contains one rotational drive pair located on the base, one translation drive pair (the axes of the two pairs overlap), and also two cardan joints, each of which is in the form of two rotational kinematic pairs with perpendicular intersecting axes arranged in a horizontal plane (Fig. 3a). Single screws characterizing the situation of the axes of these kinematic pairs have the coordinates E11(0, 0, 1, , , 0), E12(0, 0, 1, , , 0), E13(0, 0, 1, , , 0), E14(0, 0, 0, 0, 0, 1), E21(0, 0, 1, , , 0), E22(0, 0, 1, , , 0), E23(0, 0, 1, , , 0) = E13(0, 0, 1, , , 0), E24(0, 0, 0, 0, 0, 1) = E14(0, 0, 0, 0, 0, 1), E31(0, 0, 1, 0, 0, 0), E32(0, 0, 0, 0, 0, 1), E33(e33x, e33y, 0, , , ), E34(e34x, e34y, 0, , , ), E35(e35x, e35y, 0, , , ), E36(e36x, e36y, 0, , , ). Note that e33x = e35x, e33y = e35y, e34x = e36x, e34y = e36y",
+            " The first and second kinematic chains are imposed on two connections that can be considered as repetitive; they determine the number of degrees of freedom, equal to four. The third kinematic chain is not imposed on the connections. Power screws of connections due to kinematic chains, as in the previ e11x 0 e11y 0 e12x 0 e12y 0 e13x 0 e13y 0 e21x 0 e21y 0 e22x 0 e22y 0 e23x 0 e23y 0 e13x 0 e13y 0 e33x 0 e33y 0 e33z 0 e34x 0 e34y 0 e34z 0 e35x 0 e35y 0 e35z 0 e36x 0 e36y 0 e36z 0 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 39 No. 5 2010 DEVELOPMENT OF MECHANISMS OF PARALLEL STRUCTURE 411 ous cases, have the coordinates (Fig. 3b) R1(0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0). Accordingly, all the kine matic screws of motion of the output link again can be represented as screws mutual to specified power screws \u21261(0, 0, 0, 1, 0, 0), \u21262(0, 0, 0, 0, 1, 0), \u21263(0, 0, 0, 0, 0, 1), and \u21264(0, 0, 1, 0, 0, 0). Special provisions relating to the loss of one or more degrees of freedom arise if the kinematic screws corresponding to the unit vectors Ei1, Ei2, Ei3 (i = 1, 2) or to the unit vectors E33, E34, E35, E36 are linearly dependent. This occurs if any three screws Ei1, Ei2, Ei3 (i = 1, 2) or four screws E33, E34, E35, E36 fall in one plane. In particular, if any three screws Ei1, Ei2, Ei3 (i = 1, 2) fall in one plane parallel to the axis y, there exist three power screws imposed by kinematic chains R1(0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0), and R3(0, 0, 1, 0, 0, 0) (Fig. 3c) and only three kinematic screws of the movement of the output link mutual to these screws \u21261(0, 0, 0, 1, 0, 0), \u21262(0, 0, 0, 0, 0, 1), and \u21263(0, 0, 1, 0, 0, 0). Note that R3 is located on the axis y. If all drives are fixed, then, as in the previous cases, there are six force screws imposed by kinematic chains R1, R2, R3, R4, R5, and R6. This mechanism has the property of partial kinematic decoupling. Rotary engines of the first and sec ond kinematic chains move the output link in the horizontal plane"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001514_cdc.2009.5400075-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001514_cdc.2009.5400075-Figure1-1.png",
+        "caption": "Fig. 1. Gradient Projection Method. S is the feasible region, bounded by the hypersurfaces H1, H2 and G3. The supporting hyperplane of G3 at x3 is H3(x3). The projection of \u2212\u2207J(xi) onto Hi yields zi, while zd is the projection of \u2212\u2207J(x2) onto H1. Notice that to maintain feasibility at x2, it is sufficient to project onto H2. In contrast, projection onto the intersection of both active constraints corresponding to H1 and H2 will yield the zero vector.",
+        "texts": [
+            " The key mechanism that enables the method to maintain feasibility is gradient projection. Each of the k inequalities hi(x) \u2264 0, i \u2208 {1, 2, . . . , k}, defines a hypersurface Gi in R q that forms the boundary of the feasible region S = {x \u2208 R q | h(x) \u2264 0}. On each point x\u0303 of the boundary of S, each hypersurface Gi that contains x\u0303 has an associated supporting hyperplane Hi(x\u0303) that is tangent to Gi at x\u0303. The normal of Hi at the point x\u0303 is the gradient of hi(x) at x\u0303, which will point \u201caway\u201d from S. These are illustrated in Fig. 1. Similar to many optimization methods, the gradient projection method generates a sequence {xn}, the limiting point of which would be the solution to the nonlinear program (4). Consider now, the case where all hi(x) are affine functions of x [5]. Then Gi coincides with Hi, and the boundary of S are all hyperplanes. At a particular point xn that lies in the interior of S (cf. x0 in Fig. 1), the basic step is taken in a direction to decrease J(xn), ie. in the negative gradient direction, \u2212\u2207J(xn), much like the steepest descent method. When xn lies on the boundary of S (cf. x1, x2 and x3 in Fig. 1), the step is taken in a direction \u201cclosest\u201d to \u2212\u2207J(xn) while remaining within S. In this case, if \u2212\u2207J(xn) points into the interior of S, the nominal direction is taken. Otherwise, \u2212\u2207J(xn) is projected onto the intersection of the smallest set of linearly independent hyperplanes Hi that corresponds to active constraints (hi(xn) \u2265 0) that can keep xn+1 within S (cf. z1, z2 and z3 in Fig. 1). The step is then taken in this new direction, and if some active constraints are nonlinear (cf. G3 in Fig. 1), a correction is added to drive the new point xn+1 back to S. It is important to note that this smallest set of hyperplanes may exclude some active constraints, but nonetheless ensures that such exclusion will not cause further constraint violations. This case is illustrated in Fig. 1, at the point x2. Notice here that taking a step in the direction of \u2212\u2207J(x2) will violate both constraints corresponding to H1 and H2. If we project onto both of these active constraints, (in other words, the intersection of H1 and H2), the result is the zero vector, and no progress can be made. However, projecting onto H2 alone, both constraints will be satisfied, and progress can be made in the direction z2. Notice that projecting onto H1 to get zd is ineffective, since taking a step in this direction will violate the constraint corresponding to H2",
+            " For a fixed x \u2208 R q, define the following combinatorial optimization sub-problem max I\u2208J \u2016PI(x)\u2207J(x)\u2016, subject to rank (NI(x)) = |I|, NT Iac\\I (x)PI(x)\u2207J(x) \u2265 0, (5) where the norm in the objective function is the usual Euclidean norm for R q. In words, sub-problem (5) is to find a subset of Iac such that the supporting hyperplanes whose indices are in this subset are linearly independent, the projection of \u2212\u2207J(x) onto the intersection of these hyperplanes is maximal in magnitude, and when x is evolved in the resultant direction \u2212PI(x)\u2207J(x), no constraints will be violated. With reference to Fig. 1, observe that the optimal solution at the point x2 is I\u2217 = {2}, as desired. At each fixed time, let I\u2217 be the solution of subproblem (5). The generalized continuous-time gradient projection method for problem (4) that can incorporate multiple nonlinear constraints is then given by the update x\u0307 = \u2212PI\u2217(x)\u2207J(x). Similar to [7, Appendix B], the generalized scaled continuous-time gradient projection method can be obtained from the above by setting x = \u03931x\u0304, where \u03931 \u2208 R q\u00d7q is a nonsingular constant scaling matrix"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure26-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure26-1.png",
+        "caption": "Fig. 26. Hoop stresses in rivet after unloading a) w1 b) w2",
+        "texts": [
+            " For the case w2, diameter of formed rivet head reaches maximum permissible value according to riveting process manual [2]. Rivet upsetting causes metal sheet joining and filling of the rivet hole. Because of increasing fatigue life of the riveted joint, the rivet upsetting and compressive stresses distribution in the sheets around the hole should be uniform and exceed yield stress level [18,19]. Methods for FEM analysis of riveted joints of thin walled aircraft structures 961 Radial and hoop stresses distribution in the rivet after upsetting for two cases of loading is presented in Fig. 25 and Fig. 26 respectively. Compressive stresses are transmitted to sheets material around the rivet hole during the upsetting process. Contact stresses arise on mating surfaces of metal sheets (Fig. 27, 28). During riveting process (using press machine), stresses in rivet hole reach the value of 500 MPa (Fig. 27 \u2013 case w1 and w2). The value of contact stresses after unloading is equal 300 MPa (Fig. 28 \u2013 w1, w2). Jerzi Kaniowski et al. 962 Irreversible plastic deformations of sheet material around the rivet hole remain after the riveting process"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002561_ecce.2012.6342668-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002561_ecce.2012.6342668-Figure5-1.png",
+        "caption": "Fig. 5. Rotor test. Custom designed separate end ring double cage rotor with brass outer cage and copper inner cage (a), and the corresponding rotor lamination (b).",
+        "texts": [
+            " 100% 6 66 \u00d7 + = fun rl ind mPa mPamPaF (3) When the fault occurs, the energy distribution of the signal is changed in the resolution levels related to the characteristic frequency bands of the fault. Hence, the energy excess localized in the approximation signal of interest (a6 in our case) is considered as an anomaly indication in case of rotor damage. IV. EXPERIMENTAL RESULTS Experimental tests were performed on a 380V, 7.5 Hp, induction motor with a custom-made separate end ring double cage induction motor with fabricated copper bars (Fig. 5) to test the proposed method on a motor representative of large motors. The outer cage was made of brass, which is a common in large motors for obtaining high starting torque. The tests were performed for the healthy rotor and a faulty rotor after cutting 3 of 44 rotor bars. The motor load was controlled by adjusting the field voltage of a 30 Hp DC generator coupled to the motor. Commercial current sensors and a 16 bit DAQ board were used to measure the current at 960 Hz sampling with a data acquisition window of 10 seconds under speed transient conditions from 10% to 90% of the full load. The second series of tests was carried out during an acquisition time interval of 25 seconds under a speed profile from 10% to 50% then to 90% of the full load. The bar and end ring design for the double cage rotors with separate end rings (Fig. 5) are shown in Fig. 6. In fact, the separate end ring design is more common as it provides higher starting torque (higher outer cage resistance) and is less susceptible to breakage due to the independent contraction and expansion of the two cages (less thermomechanical stress at bar and end ring joint). The outer cage is usually made of high resistivity material such as brass or bronze for higher starting torque [21]-[22]. As explained in the previous section, it is possible to track over time the signature of the two fault components (1\u00b12s)f by means of wavelet multiresolution analysis"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003643_elan.201100003-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003643_elan.201100003-Figure4-1.png",
+        "caption": "Fig. 4. Water column profiles of dissolved Co from the eastern tropical Atlantic ocean determined after UV treatment (*) and without UV treatment (~). The samples were collected at N 78 40\u2019 W 248 13\u2019 (A) and N 88 1\u2019 W 298 59\u2019 (B).",
+        "texts": [
+            " Surface active substances are known to reduce the sensitivities by competitive adsorption to the electrode [20,23]. We made similar observations: the sensitivities in high dissolved organic matter seawater samples were up to 40% lower than in deep or UV digested samples. Intercalibration experiments (SAFe program, see GEOTRACES website) have additionally shown that currently available Co methods may not recover a large Co fraction in non-UV digested samples. UV digestion was also essential with this method (Figure 4). Low Co seawater was prepared by passing UV digested seawater over a pre-cleaned and conditioned Chelex-100 column [20]. Different qualities of KBrO3 (SigmaUltra grade vs. ACS grade>99.8%, Aldrich) were compared for Co impurities because at the time of the completion of this manuscript the SigmaUltra grade KBrO3 was no longer available from the manufacturer. Using the purified seawater sample, the SigmaUltra grade KBrO3 introduced no significant blank (<5 pM). At these concentrations it was difficult to say whether the residual Co resulted from Co traces in the seawater sample or a small reagent blank",
+            " The proposed method allowed a detection limit significantly below the best method with NaNO2 and DMG (<10 pM, [9]) and a similar detection limit compared to the most sensitive previously published method with Nioxime and NaNO2 (~3 pM) [23]. Using longer adsorption times the detection limit could be further reduced to sub pM values. The Co signal was linear between 0\u20137.5 nM (7 standard additions, R2 =0.9999) which included concentrations more than 10 above those found in open ocean and coastal seawater samples. The method was used to analyse Co concentrations in the tropical eastern Atlantic. Example profiles of dissolved Co determined after UV irradiation of the samples (\u201ctotal Co\u201d) are shown in Figure 4. For comparison the plot includes results of Co analyses without the UV digestion step. The difference between the two datasets could be interpreted in context of Co speciation because it shows that a large fraction of Co was present in a non-re- active form that was probably associated with strong organic complexes. The electrocatalytic reduction of diverse reactants by Co(I)(HDMG)2 suggests that a variety of other compounds that could not be tested within this study could also come into question to enhance the Co peak"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003202_csae.2012.6272949-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003202_csae.2012.6272949-Figure3-1.png",
+        "caption": "Figure 3. Coordinate systems and the forces/moments",
+        "texts": [
+            "00 \u00a92012 IEEE The flight control module which controls the attitude and altitude by sending PWM commands to the electronic speed controllers (ECS) include two important parts, i.e., an FPGA based PWM management part and an ARM based microprocessor for flight control and mission management. The entire system communicate with flight ground station via the wireless to transmit and record the data. The ground stations is used to monitor and manage the flight platform in real-time as shown in Fig. 2. III. VEHICLE DYNAMICS The coordinate systems and free body diagram for the quadrotor are shown in Fig. 3. The body frame b b b bO x y z is attached to the center of mass of the quadrotor with bx coinciding with the preferred forward direction. We use XY-Z Euler angles to model the rotation of the quadrotor in the world frame Oxyz . The rotation matrix for transforming coordinates from Body frame to World frame is given by : c c s c c s s s s c s c s c c c s s s c s s s c s c s c c R (1) Where s and c are used as the abbreviations for sin( ) and cos( ) respectively, the variables , and represent the roll, pitch and yaw angles of the helicopter in the body\u2013fixed coordinates"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000882_4243_2008_022-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000882_4243_2008_022-Figure7-1.png",
+        "caption": "Fig. 7 Cross section of a glucose-sensing layer using fluorescence quenching of oxygen as detection method. PS PMMA support, PF polyester film, I indicator layer, N nylon membrane, Glu glucose, GL gluconolactone",
+        "texts": [
+            " Typical examples are listed in Table 5. The glucose sensors are some of the most successful, both in terms of performance and on the marketplace. Almost all functions on the basis of the enzymatic action of glucose oxidase (GOx) according to Glucose + O2 \u2192 Gluconolactone + H2O2 H2O \u2013\u2013\u2013\u2192 Gluconate + H+ + H2O2 . (2) The production of protons or H2O2, or the consumption of oxygen may be monitored optically. The cross section of a typical layer for enzymatic sensing of glucose (or other enzyme substrates) is shown in Fig. 7. It is clear from this figure that response times become longer with increasing layer thickness. As with type C sensors, problematic with such sensors is potential of interference of compounds into various steps of the sensing process. A further problem with these sensors is that enzymes are also only working in the native conformation and may undergo denaturation in response to the sample, which is typically not a problem with type C sensors, as the conformation of small molecules, particularly chromophores, is much more restricted"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure5.7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure5.7-1.png",
+        "caption": "Fig. 5.7-1 Vibration model of a vehicle rearview mirror housing.",
+        "texts": [
+            " If the stiff ness matrix is not diagonal, it is said that the system has static coupling, or that the system is statically coupled. It is also possible to have both forms of coupling. In general, we can choose any coordinate system to describe the motion of a system. However, diff erent choices of coordinates will result in diff erent equations of motion. Th at is, the choice of coordinates will defi ne the type of coupling. To see this more clearly, consider a vibration model of a vehicle rearview mirror housing, as represented in Fig. 5.7-1. Th e mirror housing is modeled as a bar attached to the ground (the vehicle body) through a shaft . Let k and kt represent, respectively, the vertical bending stiff ness and torsional stiff ness of the shaft . Let the length of the bar be L. Th e mass center of the bar is located at point 1. Static Coupling. Choosing x1 and as the generalized coordinates, where x1 is the linear vertical displacement of the mass center, and is the angular displacement of the bar about the shaft . Both x1 and are measured from the bar\u2019s static equilibrium position",
+            "10-1, but the computational eff ort is considerably less. Problems 5.1 Find the natural frequency of the system shown in Fig. P5.1. Ans. 1 0= , 2 3 2 = k m 5.2 Find the natural frequencies of torsional vibration of the system shown in Fig. P5.2. Assume that the torsional stiff ness of the shaft is K and that the inertia eff ect of the shaft is negligible compared with J1 and J2. Ans. = + \u239b \u239d\u239c \u239e \u23a0\u239f K J J 1 1 1 2 5.3 Use the Lagrangian method to derive Eqs. 5.7-4, 5.7-7, and 5.7-10, which described the coordinate coupling of the system in Fig. 5.7-1. 5.4 [5.16] A triple-rod pendulum consists of three, identical pin-connected rods moving in a vertical plane, as shown in Fig. P5.4. Let each rod have length b. Derive equations of motion for small vibrations of the system. Ans. ( / ) ( / ) ( / ) ( / )( / )7 3 3 2 1 2 5 2 01 2 3 1 + + + =g b ( / ) ( / ) ( / ) ( / )( / )3 2 4 3 1 2 3 2 01 2 3 2 + + + =g b ( / ) ( / ) ( / ) ( / )( / )1 2 1 2 1 3 1 21 2 3 3 + + + =g b 0 5.5 Derive the equations of motion for small oscillations in the systems shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000662_j.jmatprotec.2007.11.038-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000662_j.jmatprotec.2007.11.038-Figure2-1.png",
+        "caption": "Fig. 2 \u2013 Straight tooth cylindrical gear generation simulation includes: push of tooth reference cylinder of hob by a which ensures tangency of tooth pitch plane of",
+        "texts": [
+            " It follows from the generation parameters that making the whole profile engages the cutter edges 28 times whereas a pinion cutter edges necessitate 116 contacts. It results from the engagement factor. Modelling tooth flanks in the CAD environment makes use of commands of turning, copying, shifting and taking away the drawn solids of the wheel and the tools. The reciprocal turning of the toothed wheel and the tool is not smooth but rather stroke-like in character. The idea of modelling is introduced in Fig. 2. The interdependence between a wheel revolution and the shift of the tool is described with Eq. (3). The designations are as follows: b tangent shift of the tool corresponding to the wheel revolution by angle \u03d5, dt turning diameter of the toothed wheel. b = dt\u03d5 360 (3) The simulation procedure in the CAD environment, written as a macrodefinition, is as follows: rotate gear wheel {revolution of generated gear} x, y {coordinates of the centre of wheel revolution} \u03d5 {value of wheel revolution angle (Fig. 2)} move tool {tool shift corresponding to a wheel revolution} n g t 334 j o u r n a l o f m a t e r i a l s p r o c e s s i 0, 0 base point of the shift b, 0 {value of the shift (Fig. 2)} copy tool {tool copy and its placement on the turning diameter of the wheel dt} 0, 0 {base point} 0, a {location of the tool copy (Fig. 2)} subtract gear wheel tool {subtraction of the dipped tool solid volume from the gear wheel model} rscript {restarting the procedure}. The kinematics of the hobbing and chiselling simulation was presented in Fig. 3. In the case of hobbing (Fig. 3a), the tool is a model of a plain milling cutter. The computer simulation of the gear generation takes place in a three-dimensional environment. A tooth profile is formed due to the turning of the gear wheel and the tool. The helix results from the hob blade moving along the wheel axis"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003239_msec2010-34282-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003239_msec2010-34282-Figure10-1.png",
+        "caption": "FIGURE 10 \u2013BENT FOAM SAMPLE AT 200 W AND 4 mm/s",
+        "texts": [
+            " The effect of the scan velocity is shown in Figure 8 for the laser power of 200 W. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/22/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use 5 Copyright \u00a9 2010 by ASME By increasing the scan velocity, the bending angle decreases because of the lowering of the energy input. In the experimented range, best results were obtained with the lower value of the scan velocity (4 mm/s): Figure 9 shows the bent sample at the end of the 150 laser passes in the case of 150 W, and Figure 10 in the case of 200 W. Lower powers lead to higher process times but they reduce sample distortions during forming. In fact, a small distortion is visible in the sample of Figure 10 in comparison with the sample of Figure 9. By increasing the power over 200 W, not only distortion but also material melting occurs as shown in Figure 11. In the figure detail, it is evident the foam degradation which was responsible for the unusual trend of the bending curve (see Figure 7). In order to obtain the numerical solution, many assumptions have been made both for the foam geometry and for the material behavior. Moreover, typical values for convection and heat flux were used to define the thermal loads"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000220_1.2424240-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000220_1.2424240-Figure2-1.png",
+        "caption": "Fig. 2 Reference \u201edotted line\u2026 and current \u201esolid line\u2026 configurations of a plane membrane",
+        "texts": [
+            " \u20203\u2021: \u201e1\u2026 Spindle shape excision of a skin cancer; \u201e2\u2026 dogar formation following the suture of the wound edges; \u201e3\u2026 proedures used to eliminate the extrusion \u201epostsuture\u2026 or to void it \u201ewhile suturing\u2026. tage. The model is implemented in the finite element code FEAP ournal of Applied Mechanics om: http://appliedmechanics.asmedigitalcollection.asme.org/ on 05/20/201 6 and used to analyze some exemplary problems. The study refers to finite strain plane homogeneous membranes subject to in-plane loading. Figure 2 shows an exemplary plane membrane and a Cartesian reference basis with unit vectors e =1,2 ,3 . The membrane is in the plane e1 ,e2 . In the reference configuration, S0 with boundary S0 and unit normal m dotted lines , the membrane is assumed to be undeformed. In the current configuration, the material points occupy the surface S with boundary S of unit normal n. The spatial position vector, x, defines the position of the material point in the current configuration and is related to the displacement vector u by u=x-p, with p the position vector in the reference configuration",
+            " The plane deformation of the membrane is described by the mapping x= f p and by the in plane deformation gradient F, and is measured locally by the material Green\u2013Lagrange strain tensor E= e e , where for , =1,2 are the strain components in the Cartesian reference system and is the tensor product. The polar decomposition, F=RU, defines the deformation gradient in terms of the right stretch tensor, U, and the rotation tensor, R. The spectral decomposition of U gives the principal stretches , U = n n , with n as the unit eigenvectors. The principal components of E are related to the principal stretches, 1 and 2, through =1/2 2 \u22121 . The vector t in Fig. 2 defines edge forces acting per unit initial length on the boundary S of the surface The membrane forces are described by the spatial tensor T= t e e , the second Piola\u2013Kirchhoff tensor = e e PKII and the first PiolaKirchhoff tensor S=s e e PKI , where , =1,2, and t , , and s are the corresponding membrane force components. The tensors are related through =JF\u22121TF\u2212T and S=JTF\u2212T, where J=det F = 1 2. In the case of isotropic elasticity, U and are coaxial. Where the membrane wrinkles, an out of plane displacement arises, u3=u3e3, and the deformation gradient becomes highly discontinuous see Sect"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002698_s1068798x10040234-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002698_s1068798x10040234-Figure3-1.png",
+        "caption": "Fig. 3. Simulation of the adaptive technological control system in ensuring specified complex parameter Cx: maxi mum tensile stress \u03c3 in the machined surface: RC, rated characteristic Cx = f(s); s is the supply.",
+        "texts": [
+            " (11), we determine the control law for Cx in turning (on the assumption that k1 remains constant under the external perturbations, while k3 varies) where Cxv = Cx_zad is the specified value of Cx; s is the initial supply, mm/turn; Cx_izm is the value of Cx calculated from Eq. (9) with the surface roughness Ra measured during machining; s(Cx_izm) is the corrected supply. Now consider the maintenance of specified Cx in final turning, when the specified value Cx_zad = 0.67, with a tolerance \u0394 = \u00b110%. For final turning, tm = 45%, Wz = 2.8 \u03bcm, Hmax = 70 \u03bcm, uch = 1.3, \u03bb = 1. The machining conditions required for the specified Cx are as follows: s = 0.15 mm/turn, v = 180 m/min, and t = 1.0 mm. This corresponds to point 1 on the rated characteristic in Fig. 3. Under the external perturbations, Cx is shifted beyond the tolerances (points 2 and 3 in Fig. 3). Pro ceeding as in the preceding examples for the case when Cx_izm \u2265 (Cx_zad + \u0394Cx_zad), we find that Analogously, we may obtain the control laws for all the other complex parameters of the surface layer (\u03a0, Cy, Cm, etc.) Thus, the results of simulation may be regarded as satisfactory, and the control laws obtained may be Cx k0 1/6k1s k1 3\u2013( )/6 v k3/6\u2013 10001/2 ,= C( )x_izm Cx_zad \u0394Cx_zad+\u2265 Cx_izm Cx_zad \u0394Cx_zad\u2013\u2264 v const= Cxv s k1 3\u2013( )/6 Cx_izm = s Cx_izm( ) Cx_zadCxv( )6/k1 3\u2013= \u23ad \u23aa \u23aa \u23aa \u23aa \u23ac \u23aa \u23aa \u23aa \u23aa \u23ab , v k3/6 1000 k0k1 ; 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003133_1350650112464323-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003133_1350650112464323-Figure1-1.png",
+        "caption": "Figure 1. Schematic of a flexible-pad thrust bearing.",
+        "texts": [
+            " Furthermore, an approximate analytical methodology is presented for the use at preliminary design stage to obtain near-optimum LCC in a thrust bearing. Department of Mechanical Engineering, Louisiana State University, Baton Rouge, LA, USA Corresponding author: MM Khonsari, Department of Mechanical Engineering, Louisiana State University, 2508 Patrick Taylor Hall, Baton Rouge, LA 70803, USA. Email: khonsari@me.lsu.edu at Monash University on December 6, 2014pij.sagepub.comDownloaded from The schematic of a flexible-pad thrust bearing is shown in Figure 1. This bearing has a split on the side of the pads. The pad surface deformation is dependent on the lubricant pressure, while the lubricant pressure itself is dependent upon the surface deformation. The pressure and deformation are thus coupled and, consequently, an appropriate multiphysics model should be considered to simultaneously solve the governing equations for both the lubricant pressure and the pad deformation. The lubricant pressure distribution can be obtained from the Reynolds equation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002764_0954406211413520-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002764_0954406211413520-Figure2-1.png",
+        "caption": "Fig. 2 Forces applied on the journal (a) and on the floating ring (b) in a hydrodynamic regime",
+        "texts": [
+            " The clearance of the internal and external films are, respectively, c1 \u00bc R2 \u2013 R1 and c2\u00bc R4 \u2013 R3. The clearance ratio is denoted by \u00bc c2/c1 and the ratio between the external and internal radius of the floating ring is denoted by \u00bc R3/R2. The inner and outer films are supposed to have the same bearing length L and the same constant lubricant viscosity m. The displacement of the journal centre O1 with respect to the floating ring centre O2 is defined in polar coordinates by the eccentricity ratio e1 and the attitude angle 1 as shown in Fig. 2(a). The displacement of the floating ring centre O2 with respect to the fixed bush centre O is defined in polar coordinates by the eccentricity e2 and the attitude angle 2 (Fig. 2(b)). Figures 2(a) and 2(b) also show, respectively, the forces applied on the journal and the floating ring. In a hydrodynamic lubrication regime, the forces applied on the journal are the constant external load W1 and the internal fluid film reaction components F\"1 and F 1 (Fig. 2(a)). The forces applied on the floating ring are the floating ring weight W2, the hydrodynamic reactions of the internal oil film F 0 \"1 \u00bc F\"1 and F 0 1 \u00bc F 1 and the external oil film hydrodynamic reactions F\"2 and F 2 . Appling Newton second law of dynamics, the following equations of motion of the journal centre and the floating ring centre are obtained e 001 e1 02 1 \u00fe e 002 e2 02 2 cos 1 2\u00f0 \u00de \u00fe 2e 02 \u00fe e2 00 2 sin 1 2\u00f0 \u00de \u00bc F\"1 M1 \u00fe W1 cos 1 M1 2e 01 0 1 \u00fe e1 00 1 \u00fe 2e 02 0 2 \u00fe e2 00 2 cos 1 2\u00f0 \u00de e 002 e2 02 2 sin 1 2\u00f0 \u00de \u00bc F 1 M1 W1 sin 1 M1 8>>>>>>>>>>>>< >>>>>>>>>>>>: \u00f01\u00de e 002 e2 02 2 cos 1 2\u00f0 \u00de\u00fe 2e 02\u00fee2 00 2 sin 1 2\u00f0 \u00de \u00bc F\"2 M2 cos 1 2\u00f0 \u00de\u00fe F 2 M2 sin 1 2\u00f0 \u00de\u00fe F\"1 M2 \u00feW2 M2 cos 1 2e 02 0 2\u00fee2 00 2 cos 1 2\u00f0 \u00de e 002 e2 02 2 sin 1 2\u00f0 \u00de \u00bc F\"2 M2 sin 1 2\u00f0 \u00de\u00fe F 2 M2 cos 1 2\u00f0 \u00de F 1 M2 W2 M2 sin 1 8>>>>>>< >>>>>>: \u00f02\u00de In the above equations, M1 is the half of the mass of the rigid rotor, M2 is the mass of the floating ring and ()0 denotes the derivative with respect to time t"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000575_taes.1972.309615-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000575_taes.1972.309615-Figure3-1.png",
+        "caption": "Fig. 3. Variance of filtered position at t i. The letters with underbars correspond to boldface symbols in the text.",
+        "texts": [
+            " Computationally, kav and krms are easily calculated from the Kalman gains already found while determining k,. A computer program was written to find k; convergence on an IBM 360/91 was obtained in two seconds from an initial estimate of a = ,B = 0.5. The k was found WILSON: CONSTANT GAIN FILTERS 839 to be unique (H concave). However, the nonlinearities in a and /B have precluded proving the uniqueness and existence of k. Whereas a Kalman filter minimizes each element of pf, by the definition of J, the constant gain filter using k minimizes the average weighted sum of the elements of ps. Fig. 3 and Table II compare the variances of the filtered position estimates as produced by each filter. There is little doubt that k produces the superior con- stant gain filter. Using k as the first-stage gain, a twostage piecewise constant gain filter was designed using kss as the second stage. When should kss be switched in? Discrete optimal control cannot provide the optimum switching times, since by definition the system controlled is not continuous in time. But, the switch should occur before the optimum trajectory intersects that of kss (tg in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003133_1350650112464323-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003133_1350650112464323-Figure2-1.png",
+        "caption": "Figure 2. Mesh and boundary conditions.",
+        "texts": [
+            " Due to the complex geometry and loading condition, it is not possible to find an analytical solution for the pad deformation. Thus, the finite element method is used to accurately predict the pad deformation. In this work a commercial finite element solver (ANSYSTM) is employed to calculate the pad deformation. In the simulations it is assumed that the pads\u2019 bottom surface is constrained from movement and the lubricant pressure is applied on the pad top surface. Since all the pads have the same geometry and experience an identical load (see Figure 2), one can take advantage of the symmetry and analyze only one single pad. An iterative scheme, as shown in Figure 3, is used to solve the multi-physics model. Based on this scheme, first an initial minimum film thickness and an initial pressure distribution is assumed and the corresponding pad deformation is evaluated. Based on the pad deflection ( ) the value of h in equation (1) is updated as follows h \u00bc 1\u00fe =h0 \u00f02\u00de where h0 is the minimum film thickness. Next, the Reynolds equation is solved for the pressure which is used as a loading function input to the FEM analysis to calculate the pad deformation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003890_s10440-011-9634-6-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003890_s10440-011-9634-6-Figure1-1.png",
+        "caption": "Fig. 1 Two points on two lines",
+        "texts": [
+            " Proposition 6 The possible rigid-body displacements allowed by a body constrained in such a way that two of its points are restricted to move on a pair of lines, forms a 2-dimensional variety with degree 6. Proof The possible displacements can be split into a rotation about the axis joining the two points followed by a displacement of the points along their lines. To understand the displacement of the points along the lines it is useful to project the lines into the plane perpendicular to the common normal to the lines, see Fig. 1. This reduces the problem to a well known problem in planar kinematics. The motion of a rigid bar so that two of its point remain on a pair of given lines is know as Cardan motion, see [1, Chap. 9 \u00a711]. This motion can be thought of as a rotation about the first point followed by a translation along the first line until the second point regains contact with its line. To summarise, any displacement which moves a pair of points in such a way that they remain on a pair of given lines can be written as the product of three dual quaternion factors. Using the coordinates given in Fig. 1 the factors can be written, g = ( 1 + \u03b5 \u03b4 2 i ) (c + sk)(c\u2032 + s \u2032p12), where p12 is the unit pure quaternion corresponding to the vector joining the two points, and c\u2032 = cos(\u03c6/2), s \u2032 = sin(\u03c6/2) with \u03c6 the rotation angle about the line. The second factor gives a rotation about the vertical, the common perpendicular to the lines with c = cos(\u03b8/2) and s = sin(\u03b8/2). After a little trigonometry, the translation along the first axis is given by, \u03b4 = 2l ( cos(\u03b1 \u2212 \u03b80) sin\u03b1 ) sc + 2l ( sin(\u03b1 \u2212 \u03b80) sin\u03b1 ) s2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure8-1.png",
+        "caption": "Fig. 8 The right-angled gearboxes (RGBs) of limb I: (a) the actual RGB and (b) a threedimensional model of the RGB",
+        "texts": [
+            " In the light of equations (39) to (42b), the twists of all the parts of the EGT are proportional to the motor speeds q\u0307i in the case of limb I and to their counterparts q\u0307i + 2 in the case of limb II, for i5 1, 2, the proportionality factor in both cases being constant. As a consequence of note 1 above, the contribution of the EGT to the product TTWMT, and, hence, its counterpart T\u0304TWMT\u0304, vanishes (see Exercise 7.8 of reference [14]). Moreover, as a consequence of notes 2 to 4, matrix T or, correspondingly, T\u0304, as pertaining to the EGT, is constant, and hence its time derivative vanishes. Therefore, the contribution of the EGT to the C-matrix is zero. The right-angled gearboxes, shown in Fig. 8(a), are installed between the drives and the proximal P joints. Their main function is the generation of the tilt motion from the rotation of the planets 6 with respect to their planet carriers 8. The gear ratio of the RGB is 1:1, and hence this subsubsystem does not have any effect on the joint kinematics of the system. However, its inertia has to be taken into account as its parts are made of steel. This subsubsection aims at deriving the contribution of the RGB to the inertia and the C-matrices of the system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure4.30-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure4.30-1.png",
+        "caption": "Figure 4.30 Vector ratio approach to frequency response",
+        "texts": [
+            " We can say from the previous text on Laplace transforms that for this specific case with a single root the length of the vector from any point on the j axis to the root on the real axis represents both the magnitude and the phase response of the system to frequency excitation. When the selected point is at the origin of the plane representing zero frequency the vector becomes a reference vector corresponding to the steady state response of the system. As we select different points p along the j axis as shown on Figure 4.30, the angle and length of the vector changes and the ratio between each of these vectors and the zero frequency (steady state) vector defines the frequency response of the system. Now, since we are working with roots of the characteristic equation (1+loop= 0 and the characteristic equation is always in the denominator of the response equation, the frequency response is the reciprocal of the response vector divided by the reference vector. Thus the amplitude ratio is the length of the reference vector divided by the response vector Root Locus 149 lengths at the various points along the j axis and the phase angle is minus the angle between the reference vector and the response vectors"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-50-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-50-1.png",
+        "caption": "Figure 3-50: Advanced abrasive wear.",
+        "texts": [
+            " The result of gross change in bearing internal geometry has been pointed out. Bearing manufacturers realize the damaging effect of dirt and take extreme precautions to deliver clean bearings. Not only assembled bearings but also parts in process are washed and cleaned. Freedom from abrasive matter is so important that some bearings are assembled in airconditioned white rooms. Dramatic examples of combined abrasive-particle and corrosive wear, both due to defective sealing, are shown in Figures 3-48 and 3-49. Figure 3-50 shows Figure 3-48: Advanced abrasive wear. Figure 3-49: Advanced abrasive wear. Machinery Component Failure Analysis 123 a deep-groove ball bearing which has operated with abrasive in it. The balls have worn to such an extent that they no longer support the cage, which has been rubbing on the lands of both rings. In addition to abrasive matter, corrosive agents should be excluded from bearings. Water, acid, and those agents that deteriorate lubricants result in corrosion. Figure 3-51 illustrates how moisture in the lubricant can rust the end of a roller"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000971_s00170-007-1176-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000971_s00170-007-1176-8-Figure1-1.png",
+        "caption": "Fig. 1 Crowning mechanism of the gear shaving machine",
+        "texts": [
+            " The model can be extended in the future into the field of controller program design in power shaving which is a newly arising technology of gear shaving for better performance and efficiency. Based on this model, this paper investigates the important influences of machine setting parameters and cutter assembly errors that would enhance the quality of shaved gears. To derive the locus equations of the shaved gear, the coordinate system has to be constructed first. The crowning mechanism of the gear shaving machine, shown in Fig. 1, can induce lead crowning on shaved gear by rocking the work table. In the motion, the pivot can be fed horizontally only, and the pin will move along the guideway. Once the angle \u03b8 between the guideway and the horizontal is specified (\u22600) in the shaving process, the rocking motion of the work table can be achieved. When \u03b8=0, the work table will move horizontally without rocking and will therefore not produce any crowning effect. The crowning mechanism can be further parameterized as shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003097_tasc.2010.2041208-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003097_tasc.2010.2041208-Figure2-1.png",
+        "caption": "Fig. 2. The computed domain of the mechanical field.",
+        "texts": [
+            " When the giant magntostrictive acceleration sensor is excited sinusoidally, it is convenient to represent the variables in a complex vector. This means that each and any of the electromagnetic and mechanic variables are represented by a phasor quantity. The mechanical field boundary should be the contour of the magntostrictive sensor. By neglecting the distortions of the rest parts of the sensor, the magntostrictive rod can be considered as the computed domain of the mechanical problem. The mechanical field boundary condition is shown in Fig. 2. axis is chosen as the direction of the exerted acceleration. , , and are the displacement components of the rod in , , and directions, respectively. The elastic energy is as follows , where and are the stress and strain tensor, respectively. The relation between stress and strain tensor is , is the elastic module matrix. The work of the external forces is given by , where and represent the external surface force density and the external volume force density, respectively, the boundary of mechanical domain , and , the displacement complex vector"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003412_978-3-642-28359-8_10-Figure10.13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003412_978-3-642-28359-8_10-Figure10.13-1.png",
+        "caption": "Fig. 10.13 Schematic illustration of methodologies adopted for LRM of (a) bimetallic wall, and (d) bimetallic tube, with the images of structures (b and e) and associated macrostructures (c and f)",
+        "texts": [
+            " The approach has been used to (1) build up bimetallic walls and tubular components and (2) to control fracture behavior of composite parts by providing compositional grading across the associated interface. A bimetallic wall and a tubular bush were fabricated by LRM with an in-house developed CW CO2 laser, coupled with a CNC work station [35\u201337]. The bimetallic wall comprises SS 316L on one side and St-21on the other side, whereas, the tubular structure consisted of St-21 on the inner side and SS 316L on the outer. Chemical composition of powders used is presented in Table 10.2. Figure 10.13 presents schematic illustrations of methodologies adopted for LRM of bimetallic wall and tubular structure, along with their photographs and associated macrostructures. Laser rapid manufacturing of bimetallic wall involved alternate deposition of two adjacent clad tracks of SS 316L and St-21, with small zone of overlap at the center, as shown in Fig. 10.13a. Two separate powder feeders, positioned on opposite sides of the incident laser beam, were used to feed SS 316L and St-21 powders during the experiment. Photograph of the bimetallic wall and its cross-sectional macrostructure are presented in Fig. 10.13b and c, respectively. Etching contrast between clad layers on opposite sides of the bimetallic wall, as seen in Fig. 10.13c, is indicative of the difference between their chemical compositions. On the other hand, LRM of bimetallic tube employed a coaxial powder feeding nozzle to deposit four partly overlapping concentric circular clad tracks in each layer. The two inner clad tracks were deposited with St-21 powder whereas the two outer clad tracks were made with SS 316L, as shown in Fig. 10.13d. Final dimensions of bimetallic tube shown in Fig. 10.13e were: 25 mm inner diameter with 3.8 mm wall thickness. Sharp etching contrast developed on the cross-section of the bimetallic tube as seen in Fig. 10.13f indicates large difference in chemical composition at the interface. In addition, an SS tube (post-machine dimensions: 34 mm ID and 2 mm wall thickness) with an internal step of St-21 (height: 1.5 mm and width: 6.5 mm) has also been fabricated by LRM, as shown in Fig. 10.14. This kind of structure will be useful for fabricating components where an insert is required to provide an internal hardfaced lining at selective places. This demonstrates the capability of LRM to add functional overhanging features for critical components"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000196_iros.2006.282213-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000196_iros.2006.282213-Figure1-1.png",
+        "caption": "Fig. 1 Error of waist orientation is accumulated by employing only",
+        "texts": [
+            " Nonlinear Compliance Control Previous landing pattern modification method is based on nonlinear virtual compliance control [15]. The foot follows uneven terrain in the last half of a swing phase, and the roll and pitch foot motion returns to the reference walking pattern in the first half of a swing phase. The reference walking pattern is set that the foot is parallel to the waist. If the waist is not parallel to the ground due to the structural deflection, the response errors of motors and so on, the foot cannot horizontally land a flat terrain (see Fig. 1). Therefore, the error of waist orientation is accumulated during walking. To solve this problem, the simplest way is to feedback the information of a gyroscope. However, it is preferable to compensate waist orientation errors with only force sensors. Because the previous landing pattern modification method is employed without any other sensors except force sensors. Furthermore, a highly accurate gyroscope is so expensive that the number of sensors should be as small as possible. So, we developed a predictive attitude compensation control using 6-axis force torque sensor"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001960_00368791011076227-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001960_00368791011076227-Figure1-1.png",
+        "caption": "Figure 1 A two-lobe four recess hybrid journal bearing system",
+        "texts": [
+            " Thus, the aim of the present research work is to bridge the gap in the literature and to numerically simulate the static and dynamic performance of a two-lobed four recessed hybrid journal bearing system compensated by capillary as well as orifice restrictors using finite element method. Further, a comparative performance of a two-lobe four recess hybrid journal bearing and a corresponding circular four recess journal bearing has also been presented for a capillary and orifice restrictors in order to have a better physical insight. The geometry of a two-lobe four recess hybrid journal bearing system has been shown in Figure 1. The generalized Reynold\u2019s equation governing the laminar flow of an isoviscous incompressible lubricant in the clearance space of a journal and bearing in non-dimensional form is expressed as (Huebner, 1975, Sinhasan et al., 1989): \u203a \u203aa h3 6 \u203ap \u203aa \u00fe \u203a \u203ab h3 6 \u203ap \u203ab \u00bc V \u203ah \u203aa \u00fe 2 \u203ah \u203a t \u00f01\u00de Study of two-lobe four recessed hybrid journal bearing Satish Jain, Satish Sharma, J. Sharana Basavaraja and Prashant Kushare Volume 62 \u00b7 Number 6 \u00b7 2010 \u00b7 332\u2013340 D ow nl oa de d by L U N D U N IV E R SI T Y A t 1 3: 19 1 8 Ju ly 2 01 6 (P T ) The expression for non-dimensional fluid-film thickness for a multilobe rigid journal bearing with reference to fixed coordinate axis is given as (Ghosh and Satish, 2003a, b): h \u00bc 1 d 2 \u00f0Xj 2X i L\u00decosa2 \u00f0Zj 2 Z i L\u00desina \u00f02\u00de where Xj and Zj are the equilibrium coordinates of the journal centre"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003017_s13369-010-0022-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003017_s13369-010-0022-8-Figure1-1.png",
+        "caption": "Fig. 1 The schematic diagram of the stabilized platform",
+        "texts": [
+            " Systems like satellite tracking antennas or critical items such as surgery tables or even billiard tables need to be isolated from the sea wave-related pitching and rolling of the ocean liners. Stabilized platforms are used as mounts for systems which are meant to be decoupled from sea waves. A two-degree-of-freedom (2DOF) parallel manipulator stabilized platform system is constructed to reject such torque disturbances and keep its top plate level with respect to the horizontal axis. The schematic diagram of the stabilized platform is shown in Fig. 1. The stabilized platform has a top-plate and a base-plate linked by two variable-length electro-mechanical actuators with the help of spherical joints. The angular motion of the top plate with respect to the base plate is produced by reducing or extending the actuators\u2019 length to reject disturbances at the base. The outputs of the stabilized platform are top-plate angular position monitored by highly precise sensors. These outputs are again fed back to a controller for error adjustment. The controller produced two separate voltage commands for each actuator"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002478_dscc2012-movic2012-8641-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002478_dscc2012-movic2012-8641-Figure3-1.png",
+        "caption": "Figure 3. STEADY STATE TURNING MANEUVER OF SKID STEERED WHEELED VEHICLE.",
+        "texts": [
+            "org/about-asme/terms-of-use rain dependent minimum turn radius (MT R) prediction curves are introduced and their applicability to path planning on sloped terrains is discussed. The remainder of the paper is organized as follows. Section 2, presents the proposed dynamic model. Section 3 includes experimental results and a description of the MT R curves. Finally, Section 4 presents concluding remarks and future work. Following [5], the dynamic model for a skid-steered vehicle performing a turning maneuver (see Fig. 3) can be expressed in terms of wheel states and is given by Mq\u0308+C(q, q\u0307)+G(q) = \u03c4, (1) where q = [\u03b8i \u03b8o] T is the angular position of the inner and outer side wheels respectively, q\u0307 = [wi wo] T is the angular velocity of the inner and outer wheels, \u03c4 = [\u03c4i \u03c4o] T is the torque of the inner and outer motors, M is the mass matrix, C(q, q\u0307) is the resistive, and G(q) is the gravitational term. Here, we concentrate on steady state maneuvers and therefore Mq\u0308 = 0. It is important to note that the proposed model considers only the slipping/skidding introduced due to steering and assumes that for linear motion pure rolling is satisfied",
+            " However, in this work, the C(q, q\u0307) model is extended to capture the frictional forces involved in the full range of possible turn radii and also the effect of terrain slope and vehicle heading. The modeling of the terrain ground interaction is based on the relationship between shear stress \u03c4ss and shear displacement j given by [3] \u03c4ss = p\u00b5(1\u2212 e\u2212 j/K), (2) where p is the normal pressure, \u00b5 is the coefficient of friction, and K is the shear deformation modulus. It is then possible to compute the longitudinal frictional forces on each side of the vehicle by integrating Eq. (2) over the contact patch of each wheel (shaded areas in Fig. 3). Assuming that the vehicle turns with constant angular velocity \u2126z about a center of turn O with turn radius R\u2032 as illustrated in Fig. 3, it is possible to compute the shear displacement for the outer and inner vehicle sides along the X and Y directions. The vehicle of Fig. 3 is assumed to have wheels of radius r, wheel base L, track width B, and wheel contact patches of size pl \u00d7 b. The contact patches were measured using the Tekscan pressure measurement system [11], which employs an array of pressure sensors placed under the tire to generate a profile of its imprint on the surface. Although the actual shape of the contact patch is elliptical, it can be closely approximated by a rectangular shape. In addition, since the tire pressure was maintained constant at a high value of 20psi, it is assumed that the size of the contact patches remain constant for all experiments. For generality sake, the vehicle is assumed to have an off centered center of gravity (CG) located at (Cx, Cy) and a center of turn shifted by an amount So from the CG. Table. 1 lists the key parameters involved in the robot model. Following [3] and the conventions of Fig. 3, the shear displacements for the inner side (front and rear) wheels at a point (xi,yi) are given by jXi f = (R\u2032\u2032\u2212 B 2 \u2212Cx + xi)cos ( (na\u2212 yi)\u2126z rwi ) \u2212 ( l 2 \u2212Cy\u2212S0)+ yi cos ( (na\u2212 yi)\u2126z rwi ) , (3) jYi f = ((R\u2032\u2032\u2212 B 2 \u2212Cx + xi)sin ( (na\u2212 yi)\u2126z rwi ) \u2212 ( l 2 \u2212Cy\u2212S0)+ yi cos ( (na\u2212 yi)\u2126z rwi ) , (4) ji f = \u221a jX2 i f + jY 2 i f , (5) where na = l 2 \u2212Cy\u2212S0, R\u2032\u2032 = R\u2032 cos\u03b2 and l = 2(L 2 + pl 2 ). jXir = (R\u2032\u2032\u2212 B 2 \u2212Cx + xi) { cos ( (nb\u2212 yi)\u2126z rwi ) \u22121 } \u2212 yi sin ( (nb\u2212 yi)\u2126z rwi ) , (6) 3 Copyright \u00a9 2012 by ASME Downloaded From: http://proceedings",
+            ", the line connecting the center of the inner wheels and the corresponding line for the outer wheels) and obtain the following normal forces on the inner (Ni) and outer (No) sides as Ni = W cos\u03b8(B 2 \u2212Cx)+Whsin\u03b8sin\u03c8\u2212 Whv2 gR\u2032 cos\u03b2 B , (21) No = W cos\u03b8(B 2 +Cx)\u2212Whsin\u03b8sin\u03c8+ Whv2 gR\u2032 cos\u03b2 B , (22) where \u03b8 and \u03c8 represent the slope of the hill and the vehicle heading, W is the vehicle weight, which as shown in Fig. 4 has a longitudinal component Wln and a lateral component Wlt , h is the height of the CG, R\u2032 is the distance from O to the CG, and \u03b2 represents the angle between the vector going from the center of turn O to the CG and a vector going from O to Ov (refer to Fig. 3 for easy visualization of this parameter). Similarly, performing a moment balance about the front and rear axles, we get N f = W cos\u03b8( l 2 +Cy)\u2212Whsin\u03b8cos\u03c8\u2212 Whv2 gR\u2032 sin\u03b2 l , (23) Nr = W cos\u03b8( l 2 \u2212Cy)+Whsin\u03b8cos\u03c8+ Whv2 gR\u2032 sin\u03b2 l , (24) Assuming symmetry, the normal pressures on each wheel are then estimated using po f = N f W cos\u03b8 No b pl , (25) por = Nr W cos\u03b8 No b pl , (26) pi f = N f W cos\u03b8 Ni b pl , (27) pir = Nr W cos\u03b8 Ni b pl . (28) The resistive torque C(q, q\u0307), has an additional component due to the rolling resistance and friction in the motor bearings and different components that make up the driving system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002032_s1793048009000946-Figure19-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002032_s1793048009000946-Figure19-1.png",
+        "caption": "Fig. 19. Cargo transport by 2 plus (blue) and 2 minus (yellow) motors: possible configurations (0), (+), and (\u2212) of motors bound to the microtubule. For configuration (0), the motors block each other so that the cargo does not move. For configuration (+) and (\u2212), the cargo exhibits fast plus and minus motion, respectively.35",
+        "texts": [
+            " Bi-directional transport by two motor species In biological cells, the motion of cargo particles along microtubules is often observed to be bi-directional in the sense that the particle frequently switches its direction of motion. Since both kinesin and dynein motors are bound to these particles, it is rather natural to assume that the bi-directional motion arises from the competition between these two motor species. The molecular mechanism underlying this competition has been controversial for some time. Two scenarios have been discussed:33, 34 (i) Tug-of-war between two motor teams: Each motor species tries to move the cargo into its own direction, thereby performing a tug-of-war on the cargo as illustrated in Fig. 19; and (ii) Coordination by a putative protein complex: Such a complex could prevent opposing motors from being active at the same time, thereby excluding state (0) in Fig. 19. The observed complexity of bidirectional transport has led many authors to reject a tug-of-war scenario and to search for a coordination complex. However, as recently shown in Ref. 35, this conclusion was premature because the stochastic nature of a realistic tug-of-war leads to rather complex transport behavior as observed experimentally. Thus, let us consider a team of plus and a team of minus motors that pull in opposite directions; the direction of instantaneous motion is determined by the stronger team as in the two states (+) and (\u2212) of Fig. 19. However, since the number of motors that actually pull varies with time in a stochastic manner for both motor species, the weaker team may suddenly become the stronger one which reverses the direction of motion. Indeed, because of the stochastic unbinding and rebinding of the motors, each individual motor experiences a strongly fluctuating load force. The instanteneous value of this force depends both on the number of kIf one takes backward steps into account, the terms with n < F/Fs give a small and negative contribution to the sum in (4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000274_bi00708a004-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000274_bi00708a004-Figure5-1.png",
+        "caption": "FIGURE 5: pH dependence in HzO of the effect of dibutyryllecithin concentration on the slopes of double reciprocal plots of velocity as a function of Caz+ concentration. The symbols refer to pH values shown in Figure 4.",
+        "texts": [
+            " Kinetic constants were evaluated by replotting slopes and intercepts from l / V us. l/[Ca] plots at different fixed levels of dibutyryllecithin. The intercept replots as a function of the fixed levels of dibutyryllecithin in HzO at various pH values are shown in Figure 4. The slope of this plot is Kpc/V, and it is seen that pH does not affect this parameter. The intercept at infinite [phosphatidylcholine] is l/ V,. This paramete1 is clearly pH dependent. The intercept at 1jV = 0 is l/Kpc, which is also pH dependent. The slope replots are shown in Figure 5. The slope of this plot is K,caKpc/V,,,. Since KFc/V,,, is pH independent, K,ca must also be pH independent. These data are also consistent with H+ inhibition occurring by an uncompetitive mechanism. In D20 data similar to those in Figures 4 and 5 were also obtained. Table I1 summarizes the kinetic constants obtained 2266 B I O C H E M I S T R Y , V O L . 1 3 , N O . 1 1 , 1 9 7 4 K I N E T I C A N D S P E C T R A L P R O P E R T I E S O F P H O S P H O L I P A S E A2 V Y Lo I- W PC as variable substrate / FIGURE 2: Inhibition caused by protonation in a bi-ter mechanism"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.16-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.16-1.png",
+        "caption": "Fig. 1.16. 'Four-wheel' steering to make slip angle fl = 0 ( Exercise 1.2).",
+        "texts": [],
+        "surrounding_texts": [
+            "90 degrees. The phase increase at low frequencies and higher speeds is due to the presence of the speed V in that same term. At speeds beyond approximately the characteristic speeA, the corresponding (last) term in the denominator has less influence on the initial slope of the phase characteristic. The lateral acceleration response (1.78) shown in the centre graph of Fig. 1.15 gives a finite amplitude at frequencies tending to infinity because of the presence of 092 in the numerator. For the same reason, the phase lag goes back to zero at large frequencies. The side slip phase response tends to -270 degrees (at larger speeds) which is due to the negative coefficient ofjo9 in the numerator of (1.79). This in contrast to that coefficient of the yaw rate response (1.77). It is of interest to see that the steady-state slip angle response, indicated in (1.79), changes sign at a certain speexl V. At low speeds where the tyre slip angles are still very small, the vehicle slip angle obviously is negative for positive steer angle (considering positive directions as adopted in Fig. 1.11). At larger velocities the tyre slip angles increase and as a result, fl changes into the positive direction. Exercise 1.2. Four-wheel steer, condition that the vehicle slip angle vanishes Consider the vehicle model of Fig.l.16. Both the front and the rear wheels can be steered. The objective is to have a vehicle moving with a slip angle,8 remaining equal to zero. In practice, this may be done to improve handling qualities of the automobile (reduces to first-order system!) and to avoid excessive side slipping motions of the rear axle in lane change manoeuvres. Adapt the equations of motion (1.46) and assess the required relationship between the steer angles ~ and 02. Do this in terms of the transfer function between c~2 and ~ and the associated differential equation. Find the steady-state ratio (02/~ )ss and plot this as a function of the speed V. Show also the frequency response function 02/~, (jco) for the amplitude and phase at a speed V= 30 m/s. Use the vehicle parameters supplied in Table 1.1. TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 37"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000182_6.2006-6609-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000182_6.2006-6609-Figure5-1.png",
+        "caption": "Figure 5. Coordinate illustrations.",
+        "texts": [
+            " In this section, a target tracking problem using visual information from a single camera fixed on the UAV, called the follower, is formulated. Without loss of generality we assume that the optical axis of the camera is aligned with the follower\u2019s longitudinal xB-axis of the body frame FB = {xB , yB , zB}, and the optical center of it is fixed at the center of gravity of the follower. Otherwise, the coordinates of the optical center and the direction of the optical axis could be taken into account in the equations of relative motion presented. Then the body frame can be chosen coincident with the camera frame; see Figure 5 (a). It is assumed that the image processing algorithm associated with the camera provides three measurements in real time. These are the pixel coordinates of the image centroid (yI , zI) in the image plane I and the image length bI in pixels; see Figure 5 (b). Assuming that the camera focal length l is known, the bearing angle \u03bb, and the elevation angle \u03d1 can be expressed via the measurements through the geometric relationships tan\u03bb = yI l , tan\u03d1 = zI\u221a l2 + y2 I . (1) In a kinematic setting the equation of motion for the flying target is given by: R\u0307 E T (t) = VE T (t), RE T (0) = RE T0 , (2) where RE T (t) = [xE T (t) yE T (t) zE T (t)]> and VE T (t) = [V E Tx (t) V E Ty (t) V E Tz (t)]> are respectively the position and velocity vectors of the target\u2019s center of gravity in some inertial frame FE = {xE , yE , zE}"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002549_s12283-010-0037-0-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002549_s12283-010-0037-0-Figure4-1.png",
+        "caption": "Fig. 4 Sketch of the experimental setup with reference direction definition",
+        "texts": [
+            " Nevertheless, the most important effects are included in the adopted layout.Fig. 1 Definition of pedal/foot angles The shoes with the pedal have been mounted on a foot shape hinged to a beam with the top extremity shaped as a shank. The lower extremity of the beam was hinged to a mechanical interface fixed over a six-component wind tunnel balance. The balance precision is in the order of 0.01 N for all the force components. This strut allowed for different angular setting of both foot and shank as sketched in Fig. 4. During the test activity the beam inclination with respect to the test chamber was set using a clinometer, while the different foot inclinations over the beam were fixed by apposite holes and pins. The uncertainty of both these two settings was in the order of a tenth of degree. A global view of the experimental setup is shown in Fig. 5. Three different shoe models (all with European size 42) were tested: \u2022 shoe A, laced up and with a very \u2018\u2018clean\u2019\u2019 and close- fitting shape (Fig. 6a) \u2022 shoe B, strap fastened (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001371_cts.2008.4543971-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001371_cts.2008.4543971-Figure3-1.png",
+        "caption": "Figure 3. The Four vVectors Involved in the Determination of a UAV\u2019s Heading h.",
+        "texts": [
+            " The proposed decentralized search algorithm expresses individual and collective goals in four weighted directional vectors which, when summed, provide the desired future heading h of the UAV as shown in Equation 1. h = wgvg + wUAV vUAV + wcvc + wmvm, (1) where vg is the goal vector, vUAV is the UAV spread vector, vc is the vector that repels the UAV from the search area boundaries, vm is the momentum vector, and the ws are their corresponding weights. New desired headings are calculated at fixed discrete moments triggered by the arrival of new sensor data. Figure 3 shows an example of how the four vectors manifest for a particular UAV with the search boundary on its left and another cooperating UAV to its right. Shaded areas in the figure represent areas that have already been searched by sensors. In actual execution, the shaded areas are composed of a series of sensor footprints, such as the ones shown in Figure 2, sequentially plotted on the map of the search area every time a new sensor reading is taken. In detail, the vector vg points in the direction of the area around the UAV that has not been visited by any other member of the group for the longest time"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001115_j.triboint.2009.11.005-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001115_j.triboint.2009.11.005-Figure8-1.png",
+        "caption": "Fig. 8. FE model (half-symmetry) of the flat o",
+        "texts": [
+            " The contact pressure distributions obtained are then used as input for the next step of the methodology presented in this paper, which consists of calculating the friction coefficient at each contact node according to its actual contact pressure value. The FE model consists of a deformable part corresponding to the rubber specimen, modeled with three dimensional linear brick hybrid elements with eight nodes, and a rigid part corresponding to the cylindrical countermaterial and the encapsulated fixing tooling, modeled with analytical rigid surfaces. Fig. 8 shows a sketch of the FE model of the tribometer test assembly. The figure shows how the FE model has been developed considering half-symmetry with respect to the longitudinal YZ plane, which is possible due to the YZ-symmetry of the rest of the loads and displacements acting in the model. The characteristics of the FE model are as follows: Boundary conditions: o Displacements (X, Y, Z) and rotations (yX, yY, yZ) are restrained at the reference node of the analytical rigid surface which simulates the cylindrical countermaterial"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure6-1.png",
+        "caption": "Figure 6. The model of the six speed automatic transmission, including all gears, bearings, clutches and housing",
+        "texts": [
+            " For example, if the contact is away from the centre of the tooth, this will generate a turning moment on the gear, which will have to be reacted by the supporting bearings. The rear planetary gear set in this example is shown in Figure 2. Within the model, the axial load is properly taken by the addition of rolling element thrust bearings (Figure 3), as shown in Figure 4. The clutch and brake parts within the system are modeled as mass, inertia and stiffness components. The completed model is shown in Figure 6, including all gears, bearings, planet carriers and clutches. The boundary conditions of the analysis are then defined. This entails specifying the input speed and torque (or power), and which clutches or brakes are locked. The convergence scheme is shown in Figure 7. The non-linear analysis can be completed in a couple of minutes. The results show the full six degree of freedom deflection of the system, the rotational speeds of and forces acting on all components, gear and bearing misalignments"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000373_50006-8-Figure6.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000373_50006-8-Figure6.1-1.png",
+        "caption": "Fig. 6.1. Typical Material Distribution in Jet Engine",
+        "texts": [
+            " They are required to exhibit combinations of high strength, good fatigue and creep resistance, good corrosion resistance, and the ability to operate at elevated temperatures for extended periods of time (i.e., metallurgical stability). 1 Their combination of elevated temperature strength and resistance to surface degradation is unmatched by other metallic materials. Superalloys are the primary materials used in the hot portions of jet turbine engines, such as the blades, vanes and combustion chambers, constituting over 50% of the engine weight. Typical applications are shown in Fig. 6.1. Superalloys are also used in other industrial applications where their high temperature strength and/or corrosion resistance is required. These applications include rocket engines, steam turbine power plants, reciprocating engines, metal processing equipment, heat treating equipment, chemical and petrochemical plants, pollution control equipment, coal gasification and liquification systems, and medical applications. 2 In general, the nickel-based alloys are used for the highest temperature applications, followed by the cobalt-based alloys and then the iron-nickel alloys"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure21-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure21-1.png",
+        "caption": "Fig. 21. Numerical model of the neighbourhood of the mushroom rivet",
+        "texts": [
+            " After riveting process simulation external loads, based on analysis of bigger part (described in point 4), could by applied to the model. Methods for FEM analysis of riveted joints of thin walled aircraft structures 957 Numerical model of upsetting process (according to [A])A The analysis was performed for the solid mushroom rivet (shank diameter 3,5 mm) joining two aluminium sheets (thickness 1,2 mm). The three-dimensional numerical model of the neighbourhood (10,5 mm wide) of a single rivet was considered (Fig. 21). Dimensions of the mushroom rivets were taken according to Russian standard [OST 1 34040-79] (head radius R = 4,2 mm, diameter D = 7 mm, height h = 1.88 mm, without compensator). The radius Rp of the rounded tool surface was equal to 4,8 mm. Rivet and aluminium sheets were described using three-dimensional elements type Hex8 [16]. Elastic-plastic material model was considered. Simplified stress-strain curves for alloys PA25 (rivets) and D16TN (sheets) obtained from uniaxial tension and compression tests are presented in Fig",
+            " Iterative penetration checking approach allows to change the contact condition within the Newton-Raphson iteration loop. Using this procedure, the iteration process is done simultaneously to satisfy both the contact constraints and global equilibrium using the Newton-Raphson method. This procedure is accurate and stable but may require additional iterations. Numerical analysis The numerical calculations were performed for two cases of upsetting (w1, w2) distinguished by the height of the formed rivet head (Table I, Fig. 21). Methods for FEM analysis of riveted joints of thin walled aircraft structures 959 Riveted joint deformations are shown in Fig. 23, while rivet deformations for two cases of upsetting are given in Fig. 24. In the first case (w1), the height of formed rivet head has maximum value, when the diameter reaches nominal value [17]. For the case w2, diameter of formed rivet head reaches maximum permissible value according to riveting process manual [2]. Rivet upsetting causes metal sheet joining and filling of the rivet hole"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure3-1.png",
+        "caption": "Fig. 3 Generation of a dual cutter cycloid",
+        "texts": [
+            " 6 and 7 simultaneously Re sj jk, Hk, G = Med G \u00b7 Mdc G \u00b7 Mcb \u00b7 Rb sj jk, Hk 6 f jk, Hk, G = ne \u00b7 Ve HG = A sin G + Hk + B cos G + Hk = 0 7 here Mcb = 1 0 0 Amk 0 1 0 0 0 0 1 Ank 0 0 0 1 k = l,r Mdc G = 1 0 0 rm G 0 1 0 \u2212 rm 0 0 1 0 0 0 0 1 Med G = cos G + sin G + 0 0 \u2212 sin G + cos G + 0 0 0 0 1 0 = Hk = rn rm A = \u2212 nxe xc + rm G \u2212 nyeyc B = nxeyc \u2212 nye xc + rm G The symbols xc, yc, and zc are the components of the cuttersurface position vector represented in the coordinate system Sc. The symbols nxe, nye, and nze are the components of the unit normal vector outward to the cutter surface. The parameter is the ratio of the roll angle between the cutter and the work gear, while The parameter rn indicates the radius of the rolling circle. Figure 3 illustrates the machine setup of the head cutter. The pa- MARCH 2010, Vol. 132 / 031008-3 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use r t t c c r w w v c a T g d T t o 0 Downloaded Fr ameters Amk and Ank are the setup distance on the cutter swivel able with respect to machine center Oc. To ensure symmetrical ooth traces, the right and left cutters are also symmetrically loated with respect to Oc. The coordinate transformation from the oordinate system S1 to the coordinate system S2 can thus be epresented as follows: R2 = M21 \u00b7 R1 = cos Hk + k 0 sin Hk + k Amk + s 0 1 0 0 sin Hk + k 0 cos Hk + k Ank 0 0 0 1 \u00b7 rc 0 0 1 = s + Amk + rc cos Hk + k 0 Ank rc sin Hk + k 1 k = l,r 8 here s = rn Hk, Amk = rc + Wm 9 here the parameter Wm is the modified offset, and the position ector R2 denotes the equation of the cutter\u2019s cycloid"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.4-1.png",
+        "caption": "Fig. 1.4. Vehicle model showing three degrees of freedom: lateral, yaw and roll.",
+        "texts": [
+            " Chiesa and Rinonapoli (1967) were among the first to employ effective axle characteristics or 'working curves' as these were referred to by them. V~gstedt (1995) determined these curves experimentally. Before assessing the complete non-linear effective axle characteristics we will first direct our attention to the derivation of the effective cornering stiffnesses which are used in the simple linear two-wheel model. For these to be determined, a more comprehensive vehicle model has to be defined. Figure 1.4 depicts a vehicle model with three degrees of freedom. The forward velocity u may be kept constant. As motion variables we define the lateral velocity v of reference point A, the yaw velocity r and the roll angle (p. A moving axes system (A,x,y,z) has been introduced. The x axis points forwards and lies both in the ground plane and in the plane normal to the ground that passes through the so-called roll axis. The y axis points to the right and the z axis points downwards. This latter axis passes through the centre of gravity when the the roll angle is equal to zero",
+            "16) together yields the total steer angle for each of the wheels. The effective cornering stiffness of an axle Ceff, i is now defined as the ratio of the axle side force and the virtual slip angle. This angle is defined as the angle between the direction of motion of the centre of the axle i (actually at road level) when the vehicle velocity would be very low and approaches zero (then also Fyi ~0) and the direction of motion at the actual speed considered. The virtual slip angle of the front axle has been indicated in Fig. 1.4 and is designated as aa~. We have in general: F i Ceff, i - (1 .17 ) o~. at The axle side forces in the steady-state turn can be derived by considering the lateral force and moment equilibrium of the vehicle: 1 - a - ' m a (1 18) F y i - 1 Y The axle side force is the sum of the left and right individual tyre side forces. We have FyiL - (1//2CFai + ( a i Z J F z i ) (~ - ~btio) + (I//2CFTi + ~ i z J F i ) (~i - ~io ) (1.19) F i R - (1//2CFai - ( a i A V z i ) (o~i + ~llio) + (l//2CFyi - ~ i A F z i ) (~i + ~io ) where the average wheel slip angle a i indicated in the figure is O~ i - 6ta i + ~//i ( 1 ",
+            " However, the theory may approximately hold also for quasi-steady-state situations for instance at moderate braking or driving. The influence of the fore-and-aft force Fx on the tyre or axle cornering force vs slip angle characteristic (Fy, a) may then be regarded (cf. Fig. 1.9). The forces Fy~ and Fxt and the moment Mzl are defined to act upon the single front wheel and similarly we define Fy2 etc. for the rear wheel. In this section, the differential equations for the three degree of freeAom vehicle model of Fig. 1.4 will be derived. In first instance, the fore and aft motion will also be left free to vary. The resulting set of equations of motion may be of interest for the reader to further study the vehicle's dynamic response at somewhat higher frequencies where the roll dynamics of the vehicle body may become of importance. From these equations, the equations for the simple twodegree of freedom model of Fig. 1.9 used in the subsequent section can be easily assessed. In Subsection 1.3.6 the equations for the car with trailer will be established",
+            " The longitudinal forces are either given as a result of brake effort or imposed propulsion torque or they depend on the wheel longitudinal slip which follows from the wheel speed of revolution requiring four additional wheel rotational degrees of freedom. The first equation (1.34a) may be used to compute the propulsion force needed to keep the forward speed constant. The vertical loads and more specifically the load transfer can be obtained by considering the moment equilibrium of the front and rear axle about the respective roll centres. For this, the roll moments M~i (cf. Fig. 1.4) resulting from suspension springs and dampers as appear in Eq.(1.34d) through the terms with subscript 1 and 2 respectively, and the axle side forces appearing in Eq.(1.34b) are to be regarded. For a linear model the load transfer can be neglected if initial (left/right opposite) wheel angles are disregarded. We have at steady-state (effect of damping vanishes): -c~~176 + F i h i (1.35) ZJ Fzi = 2 s . l The front and rear slip angles follow from the lateral velocities of the wheel axles and the wheel steer angles with respect to the moving longitudinal x axis",
+            "34) may be further linearised by assuming that all the deviations from the rectilinear motion are small. This allows the neglection of all products of variable quantities which vanish when the vehicle moves straight ahead. The side forces and moments are then written as in Eq.(1.5) with the subscripts i = 1 or 2 provided. If the moment due to camber is neglected and the pneumatic trail is introduced in the aligning torque we have: F i - F a i + F ~ i - CFaiO~i + CFyi ~i (1.37) Mzi - Mza i - -CMaitZi - - t i F a g - - t i fFai tZi The three linear equations of motion for the system of Fig. 1.4 with the forward speed u kept constant finally turn out to read expressed solely in terms of the three motion variables v, r and ~\" m ( f + u r + h ' ~ ) - C F a l { ( 1 + C s c l ) ( u 5 + e ~ - v - a r ) / u +Csrl~O } + +CF~ 2 {(1 +Cscz) ( -v+br) /u +Csr2~ } + (CFyl't'l +CFy2g2)(/9 (1.38a) I z?+( I zOr - I x z ) (~ - (a - t l )CFa l { (1 +Cscl)(Ua + e a - v - a r ) / u +Csr,e } + - (b+ti) CFa 2 {(1 +Csc2) ( - v + b r ) /u + Csr2( P } + (a CF?lVl - b CFy2\"c2) ~ (1.38b) ( I + mh'2)i~ + mh ' (9 + u r) + ( IO r - I z ) ~ + (k~01 + k~02)~b + + (C~o 1 + Cqo 2- mgh')(p - 0 (1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001404_iembs.2009.5333492-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001404_iembs.2009.5333492-Figure3-1.png",
+        "caption": "Figure 3. Stem Cell Harvest System",
+        "texts": [
+            " While these tenets may seem like barriers, they must be recognized as the current standard of care. Design education and the practice of design thinking employs the skills necessary to problem solve through the exploration of form, among other methods. This problem solving through form and geometry is necessary when designing tools that interact with tissue, especially in cases where the device use requires intricate interaction through complex anatomy. By capitalizing on disciplinary strengths a truly collaborative solution can be accomplished. Figure 3 illustrate the collective efforts of a multi-disciplinary team from students in the University of Cincinnati Medical Device innovation and Entrepreneurship Program (UC MDIEP) student team. These sketches of a stem cell harvest device both analyze the problem visually through sketching and begin to identify potential solutions. Regulatory pathways are often viewed as a significant hurdle if best practices are not a priority for a development team. Additionally it is important to recognize that not all regulating entities have the exact same processes, expectations and standards e"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000017_6.2005-6002-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000017_6.2005-6002-Figure2-1.png",
+        "caption": "Figure 2. Le",
+        "texts": [
+            " Various implementations of the EKF are possible, employing different states 6,8, and could be augmented with an adaptive element for robust estimation in the presence of unmodeled dynamics and disturbances 9. The Image Processing block provides the position and size of the leader aircraft in the camera images. From the position, a unit-vector to the leader aircraft is computed, and the maximum angle subtended by the leader aircraft on the image plane losu\u0302 \u03b1 is computed from the size of the leader aircraft (Figure 2). The EKF model then includes the states T loslos b R R R uu \u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 ,,1,\u02c6,\u02c6 & & driven by the American Institute of Aeronautics and Astronautics 4 measurements where are the range and range-rate respectively, and b is the wing-span of the leader aircraft [ T losu \u03b1,\u02c6 ] RR &, 8. At this stage we have not integrated our adaptiv estimation. Presently, we are using the true valu ( X )\u03bb of the leader with respect to the follower. An estimates of and respectively, and an XY\u03bb& X\u03bb& relationships between range and the LOS angles w ( FL XXR \u2212= a X = ta\u03bb \u239c \u239c \u239c \u239d \u239b \u2212= \u22121tanXY\u03bb III"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001982_s12206-009-0346-z-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001982_s12206-009-0346-z-Figure2-1.png",
+        "caption": "Fig. 2. a typical 3-stage gearbox design with planetary gears.",
+        "texts": [
+            " In addition, the seamless integration and the preserved associativity within LMS Virtual.Lab avoid timeconsuming and error-prone transfer of data and allow development team to perform fast and multi-target optimization loops. In a standard 1.5MW wind turbine the huge input torque coming from the blades has to be transferred to realize a gear ratio between input and output shaft of more than 100 in order to match the needed rotational speed to generate electricity from the generator. Typically this is done through a 3-stage gearbox design (Fig. 2) with planetary gears offering a large gear ratio and a good load distribution. The gearbox model includes gears, shafts, bearings and the housing. At this level a dynamic simulation is computed and the results can be visualized through 3D animations or 2D graphs from any variable in the model. Various alternatives of the design are compared in order to optimize the system with regards to any specific performance attribute. This allows an in depth understanding of the root causes of its behavior and enables engineers to minimize the risk of failure during subsequent assessment test on prototypes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003418_jae-2012-1550-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003418_jae-2012-1550-Figure1-1.png",
+        "caption": "Fig. 1. Structure of a proposed hybrid excited FSPM motor.",
+        "texts": [
+            " First of all, the structures and operation principle of hybrid excited FSPM motor that can flux-control are explained by installing field winding on a stator. 2D FE-analysis model considering overhangs of the motor is suggested to calculate the parameters of the suggested motor. Then characteristic analysis using voltage equation is carried out in order to verify flux control characteristics and it was confirmed that operation with a high torque and speed can be achieved using flux switching. A Fig. 1(a) shows the structure of the suggested motor to which flux control characteristics are added to conventional FSPM motor. Figure 1(b) demonstrates 3D structure of core part of the stator in order to show field winding and armature winding. A permanent magnet is inserted between U-type core. Field winding is coiled as if it is surrounding the permanent magnet and the armature winding is coiled by crossing the top part. Figure 2 shows the principle of flux switching of hybrid excited FSPM motor. As shown in the figure, the magnetic flux of the gap can be controlled by controlling the direction of a field current. Figure 3 demonstrates the stator core and permanent magnet of the suggested motor"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000763_icicta.2008.153-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000763_icicta.2008.153-Figure1-1.png",
+        "caption": "Figure 1. The Sketch of active heave compensation hydraulic winch system",
+        "texts": [
+            " So the active heave compensation hydraulic winch system is adopted widely. The hydraulic winch is the traction device of the heave compensation system. In addition, an inertial measurement unit connected to a digital signal processing computer measures the heave motion of the supporting ship. The control unit determines the speed of the hydraulic winch by managing the input voltage of an electronic-hydraulic proportional valve. The sketch of an active heave compensation hydraulic winch system is shown in figure 1. In figure 1, W(t), S(t) and V(t) denote the motion of hydraulic winch, the motion of supporting ship and the motion of TMS respectively. If the length change of umbilical is ignored, then the heave motion of the TMS equals the heave motion of the ship plus the motion of hydraulic winch, described as follows: S(t) + W(t) = V(t) (1) The active heave compensation hydraulic winch is effective as long as the heave movement of the TMS is compensated to be less than the heave movement of the ship by controlling the motion of hydraulic winch"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure8-1.png",
+        "caption": "Figure 8. Geometry of Rmin path.",
+        "texts": [
+            " Then, the goal of two circles of maximum radius is to find, first, a circle which passes through the mid point M and is tangent to the current direction line, and second, a circle which passes the mid-point M and is tangent to the target direction line, as shown in Figure 7. Only a small amount of geometric manipulation is required to derive the equation for R4 as follows; (7) The case of \u03b8<0 when the target position is in the first quadrant is reserved for the reader\u2019s preference. Case 3: Employment of Minimum Radius Path If the radius of an arc path obtained in either Case 1 or Case 2 is smaller than the Minimum Radius, denoted by Rmin the minimum radius a vehicle can follow, (normally about 4.5 meters), an additional arc should be added to the Rmin arc path. Figure 8 presents a case where the target position is quite close to the Yw axis and the directional difference \u03b8 value is large. The radius of a circle that passes through a target position and is tangent to the target direction line as well as to the current direction line, i.e., Yw (dashed circle), is smaller than the value of Rmin as shown by the solid circle. To generate a path from current position C(0,0,0) to target position T(a,b,\u03b8), and Rmin circle is drawn such that it passes through the target position and is tangent to the target direction line. Then another arc path of radius R5 is added to form an S-path as shown in Figure 8, where the arc of radius R5 is a portion of a circle that passes through the current position C(0,0,0) and is tangent to the current direction line as well as to the Rmin arc. The equation for R5 is shown in Equation 8, where, its derivation steps have been omitted here for brevity of the paper. (8) From one image frame, a travel path is generated as an arcpath or S-path, and using the radius of an arc or S-path, the steering angle is determined using the bicycle model as shown in Equation (9). (9) where l indicates the length between the front wheel and rear wheel and R is the radius of the arc that the vehicle should follow"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001499_med.2009.5164579-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001499_med.2009.5164579-Figure2-1.png",
+        "caption": "Fig. 2. Formation vector model.",
+        "texts": [
+            " if (gbestV > 0 and V (pbestp) == 0) or (gbestV == 0 and V (pbestp) == 0 and F (pbestp) < gbestF ) or (gbestV > 0 and V (pbestp) > 0 and V (pbestp) < gbestV ) then gbest = pbestp gbestV = V (pbestp); gbestF = F (pbestp) end if Apply Simulated Annealing to gbest: newgbest = gbest for l = 1 to L do Move along each direction and compute pa: newxq = newgbest + \u03b7N(0, 1)(Xq \u2212 Xq) if V (newx) == 0 then if V (newgbest) == 0 then pa = min(1, exp(F (newgbest)\u2212F (newx) t )) else pa = 1 end if else if V (newgbest) == 0 then pa = 0 else pa = min(1, exp(V (newgbest)\u2212V (newx) t )) end if end if if pa >= U(0, 1) then newgbest = newx end if end for gbest = newgbest; t = \u03bbt gbestF =F (newgbest); gbestV =V (newgbest) end for return gbest Algorithm 3: Hybrid particle swarm optimization. A localization system, composed by a GPS integrated with proprioceptive sensors, is proposed which allows to determine the configuration q(t). As a result, the state d(t), that is the relative displacement of the leader with respect to the body frame projected in the horizontal plane (see Figure 2), can be considered fully accessible. The vehicle is controlled by control vector u(t) while the leader moves with velocity vector u\u2032(t) that represents an interaction term for the follower. Let sample the system with a sampling interval T and define the sampled variables qk , q(kT ), dk , d(kT ), uk , Tu(kT ), u\u2032 k , Tu\u2032(kT ). The new control vector uk is formed by finite movements performed within each sampling interval. By standard manipulation [16], the following discretetime model for the prediction of the formation vector is stated: d\u0302k+1|k = A(u\u0302k|k)d\u0302k|k + B(u\u0302k|k)u\u0302k|k + E(d\u0302k|k, u\u0302k|k)u\u0302\u2032 k|k, (7a) xk+1 \u2208 X, uk \u2208 U, vk \u2208 V (7b) where matrices A, B and E depend on yaw velocity of the vehicle and on relative orientation of the leader with respect to the considered vehicle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001369_icicic.2007.562-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001369_icicic.2007.562-Figure1-1.png",
+        "caption": "Figure 1. Configuration, free-body and transition manoeuvre diagrams of new ducted-fan VTOL UAV.",
+        "texts": [
+            " It uses a fivebladed fixed pitch propeller enclosed in an annular wing for thrust production. Geometrically, this new ductedfan UAV design is about 0.3 m in both width and length, and has approximately 3 kg in maximum takeoff weight. This vehicle will take off and land vertically and will be able to hover. Once the UAV is in vertical flight, it can transit 90o into the horizontal flight, making its fuselage parallel to the horizon. The configuration layout and the transition manoeuvre snapshot of this UAV are depicted in Figure 1(a) and (b). This UAV is currently modeled by using aerodynamic and propulsive data that has been estimated theoretically [7]. The UAV model for transition manoeuvre from vertical to horizontal flights is governed by a set of longitudinal equations of motion written in the body 0-7695-2882-1/07 $25.00 \u00a92007 IEEE axes system. These equations of motions are based on aircraft model presented in [1]. Figure 1(c) shows the diagrammatic of the UAV in body axes representation which defines all the variables of interest. The longitudinal equations of motion including the kinematic and navigational, with all zero terms (Iyz = Iyx = 0; UAV is symmetric in xz plane) dropped, are given by equations (1) to (6). u\u0307 = Fx/m\u2212 qw \u2212 g sin \u03b8 (1) w\u0307 = Fz/m + qu + g cos \u03b8 (2) q\u0307 = M/Iy (3) \u03b8\u0307 = q (4) \u02d9PN = u cos \u03b8 + w sin \u03b8 (5) h\u0307 = u sin \u03b8 \u2212 w cos \u03b8 (6) CD = CDo + [CL \u2212 CLo ]2/\u03c0.e.AR + C\u03b4 D\u03b4 (7) CL = CLo + C\u03b1 L\u03b1 + C\u03b4 L\u03b4 + c 2Va [C\u03b1\u0307 L\u03b1\u0307 + Cq Lq] (8) Cm = Cmo + C\u03b1 m\u03b1 + C\u03b4 m\u03b4 + c 2Va [C\u03b1\u0307 m\u03b1\u0307 + Cq mq](9) Equations (1) to (6) includes UAV forces and moments due to the gravitational, aerodynamic and propulsive effects"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003016_itab.2010.5687664-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003016_itab.2010.5687664-Figure3-1.png",
+        "caption": "Fig. 3. The model of the MR-compatible robotic manipulator that is currently constructed in our laboratory for performing MRI -guided transapical procedures in the beating heart. The manipulator consists of two units. The external or \"Thoracic Unit'\" facilitates access from the thorax to the apex of the heart and carries (B) and provides actuation to the Intracardiac Unit (C). The base structure (Base) in (A), is only used for dynamic simulations of the model and in the final system will be compacted for fitting into a 70 cm wide MR gantry.",
+        "texts": [
+            " The steering point for any time frame was then assigned as the intersection of the aortic annulus midline (known from the previous step) with the access corridor and the first-to-cross short axis slice. Using this computational core and its augmented reality interface, we have simulated and analyzed numerous scenarios of transapical procedures extracting dynamic trajectories [8] and investigating the relative motion of tissue with tracking[18]. Based on this analysis we have completed the virtual prototyping (on Solidworks\u00ae) and the kinematic and dynamic analysis (on Simulink and SimMechanics\u00ae) of a novel MR compatible robotic device shown in Fig. 3 for transapical interventions on the beating heart. Compared to previous works in this area, the proposed CPS introduces certain innovations with primary one being the integration of the imaging scanner and the robot into a system, thereby aIIowing on-the-tly control of both the scanner and the robot. In tum this aIIows that the same control algorithm controls the sensor (to receive exactly information it needs for its processing) and the manipulator (based on those information). The specific robotic device (Fig. 3) we develop will allow access to the beating heart without restricting its motion as previous systems [15], while alleviates the limitations of manual manipulations [16]. Currently, our effort is focused on: (1) completing the construction and bench-top testing of the robotic device shown in Fig. 3, and (2) the interfacing the computational core (Fig. 1) to an MR scanner (Siemens 1.5T Espree) for on-the-tly transfer of raw MR data for processing and for updating the imaging parameters from the core. The versatility of MRI, due to both its inherent properties and its programmability (reviewed in Tables I and II) makes this modality a strong candidate for preoperative planning and intraoperative guidance of minimaIIy invasive cardiac surgeries. From the engineering perspective, there is an ever growing body of enabling technologies that underscores that the modality can practicaIIy address current and future needs in this field"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002045_iciea.2009.5138378-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002045_iciea.2009.5138378-Figure2-1.png",
+        "caption": "Figure 2 Virtual structure geometry",
+        "texts": [
+            " CONTROL ARCHITECTURE VIA VIRTUAL STRUCTURE APPROACH In the virtual structure approach, the entire formation can be thought of as a virtual rigid body called virtual structure with formation frame FF located at its virtual center of mass. The desired states for every spacecraft in the formation can be specified in the virtual structure frame FF by the supervisor. As the formation frame FF moves, the spacecrafts track the desired positions and attitudes and the entire formation behaves as a rigid body. Figure 2 shows the coordinate frame geometry for the virtual structure approach. Denote IF as the inertial frame. Since the formation can be treated as a rigid body with inertial position Fr , velocity Fv , attitude Fq and angular velocity F\u03c9 , the formation frame FF is located at Fr with orientation given by Fq relative to IF . To describe the motion easily, the body frame of the thi spacecraft can be denoted as iF and this spacecraft can be represented either by position ir , velocity iv , attitude iq and angular velocity i\u03c9 relative to IF or by iFr , iFv , iFq and iF\u03c9 relative to the frame FF ",
+            "With these foregoing defined notations, the desired motion for every spacecraft are given by ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ( )) ( ) ( ) ( ) ( ) ( ) ( ) d d i F IF iF d d d i F IF iF F IF iF d d i F iF d d i F IF iF t t t t t t t t t t t t t t = + = + + \u00d7 = \u2297 = + r r C r v v C v \u03c9 C r q q q \u03c9 \u03c9 C \u03c9 (9) where IFC is the rotation matrix relating frame FF to frame IF , superscript d above a vector means a desired state for every spacecraft. This virtual structure foregoing described can be found the similar definition in Ref [9-12], which is suitable for the formation including small number of spacecraft. If we consider tens of or even more spacecrafts, this virtual structure will become a bit chaos that isn\u2019t suitable for mission control and operation. Focusing on this shortage, we can define multilayer virtual structure based on the Fig. 2 as: The entire formation is considered as one big virtual structure with all the foregoing properties presented in Fig.2 and the single spacecraft shown in Fig.2 is considered as the sub-VS, which includes several sub sub-VS. This style of definition will iterate until ending the single spacecraft. Compared with the traditional virtual structure, this multi-layer virtual structure has the following advantages: the scheme of this structure is very simple which is able to delivery the command to every single spacecraft quickly, particularly in the large number of spacecraft; this structure compromises the advantages of the centralized architecture and decentralized architecture, which not only has the ability to position single spacecraft rapidly to save the communication link but also ensure the autonomy of every spacecraft that will not be controlled by the center reference spacecraft absolutely and even without reference spacecraft in the mission"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000485_j.fss.2008.03.021-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000485_j.fss.2008.03.021-Figure7-1.png",
+        "caption": "Fig. 7. Two local spline basis functions.",
+        "texts": [
+            " This simple examples serves to clarify the qualitative behavior of the various methods. The generic system is x\u0307 = w + d(x, t) + u, (39) where disturbance d(x, t) incorporates approximation error and external signals. The simulation uses known ideal weights and known disturbance, rather than a specific nonlinear function. The performance then becomes independent of the type of nonlinearity, better resolving the effects of the disturbance. The approximator contains local spline membership functions (Fig. 7). For investigating the effect of oscillation of the state about a cell boundary, the inputs and parameters were chosen w = [\u22121 1]T, = 1, d = 0.1 sin(2 t), G = 1. Applying control u = \u2212 w\u0302 \u2212 Gx, (40) results in the differential equation for the state, x\u0307 = w\u0303 + d \u2212 Gx. (41) The differential equation for the ith weight error (i = 1, 2) is different depending on the method used: leakage: \u02d9\u0303wi = B( i ) [\u2212 ix + (wi \u2212 w\u0303i)], e-modification: \u02d9\u0303wi = B( i ) [\u2212 ix + |x|(wi \u2212 w\u0303i)], deadzone: \u02d9\u0303wi = B( i ) { \u2212 ix if |x| > , 0 otherwise, new method: \u02d9\u0303wi = B( i ) (\u2212 ix + i (a\u0303 \u2212 c\u0303) + (p\u0303i \u2212 w\u0303i)), \u02d9\u0303pi = B( i ) ( i (c\u0303 \u2212 a\u0303) + (\u2212o\u0302i \u2212 p\u0303i))"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001502_cdc.2008.4739357-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001502_cdc.2008.4739357-Figure4-1.png",
+        "caption": "Fig. 4. The geometric construction behind Proposition 1.",
+        "texts": [
+            " Such a lower bound has cost: tlow = ||q2 \u2212 p0||2 Vc \u2212 a\u0304 ( Vc + Vv 2Vc ) + a\u0304 ( Vc + Vv 2Vv ) (14) In many significant cases it is possible to find a geometric solution whose cost is equal to those two lower bounds, and then optimal. The main idea, underling the geometric constructions that we will present here, is to take advantage of the possibility of the carrier to launch the carried vehicle (that will visit q1) and to take a shortcut in order to have a rendezvous somewhere on the straight line to q2 before the carried vehicle has exhausted the fuel. This idea is depicted in Fig. 4. The following propositions can be stated: Proposition 1: Let \u03b80\u030212 be the smallest angle formed by the segments p0q1 and q1q2. If \u03b80\u030212 \u2264 acos ( 1 \u2212 2(V 2 c /V 2 v ) ) then it exists at least one optimal solution to the problem such that the total cost is equal to lower bound (13) Proof. Let us proceed with a constructive proof. As shown in Fig. 4, the idea is that the carried vehicle will be released at distance Vva1 from q1 on the line segment p0q1 and will be recovered at distance Vva2 from q1 on the straight path q1q2. In order to satisfy the hypothesis of the lower bound (13) and the operativeness constraints, a1 + a2 = a\u0304 (15) has to be imposed. As it is clear from the Fig. 4, between the take-off and the landing events, the carrier will \u201ccut the edge\u201d by following the straight segment qto1, ql1 whose length will be denoted by Vctcarrier. Because of operativeness constraints, this proposed solution will be a feasible one to the given problem if tcarrier \u2264 a\u0304. (16) By simple triangle consideration it is possible to see that Vctcarrier \u2264 Vva1 + Vva2 = Vv a\u0304. In particular by using Carnot theorem it is possible to write that (Vctcarrier) 2 = (Vva1) 2 + (Vva2) 2 \u2212 2V 2 v a1a2 cos ( \u03b80\u030212 ) "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000702_cca.2008.4629609-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000702_cca.2008.4629609-Figure2-1.png",
+        "caption": "Fig. 2. electronic throttle valve",
+        "texts": [
+            " In this paper two substantially different sliding-mode controllers based on a simplified second order plant model are presented. The required angular velocity of the valve plate is obtained by numerical differentiation of the opening angle. The paper is organized as follows: In Section II the mathematical model of a throttle valve is presented. Section III-A outlines the design of an integrating sliding-mode controller. In Section III-B the so-called super-twisting algorithm is applied to the throttle valve. Section IV presents simulations and experimental results, Section V concludes the paper. In Fig. 2 an electronic throttle valve is shown. It consists of a dc-motor, a gear unit, a spring mechanism and the revolving valve plate. The opening angle \u03d5 of the valve is measured by means of two potentiometers. The goal of the design is to make the angle \u03d5 track a given (sufficiently smooth) reference angle r. A strongly simplified mathematical model which describes the movement of the valve plate as a function of the motor voltage u reads as d2\u03d5 dt2 = \u2212\u03b1(\u03d5 \u2212 \u03d50) \u2212 \u03b2 d\u03d5 dt \u2212 \u03b3(\u00b5 + \u03b4 sign d\u03d5 dt ) + \u03b3 u (1) The positive parameters \u03b1, \u03b2, \u03b3, \u03b4 and \u00b5 are physical parameters of the plant which can be determined by experiment [5], \u03d50 is the so called \u201dlimp home\u201d angle",
+            " Simulation results of the discussed integrating slidingmode concept are shown in Fig. 4, where the tracking error x1 and the actuating signal u are plotted. The tracking performance of the simulated feedback loop is very satisfying with a maximum angle error of less than 0.5\u25e6. The corresponding control signal u is almost chatter-free. Controller parameters, which result from simulation provide a basis for real experiments. To improve the experimental results, the parameters were modified marginally. Results with the throttle valve in Fig. 2 are shown in Fig. 5. The feedback loop shows excellent tracking behaviour. Like in the simulation, the absolute value of the tracking error remains below 0.5\u25e6 during the whole experiment. The chattering of the control signal, which is partially due to the numerical differentiation of the measured angle \u03d5, is acceptable. Parameters of the super-twisting algorithm, which were used to produce Fig. 7, are listed in Table III. The simulation results are comparable to those obtained with the standard sliding-mode controller, see Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001219_s12541-009-0102-4-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001219_s12541-009-0102-4-Figure1-1.png",
+        "caption": "Fig. 1 Schematic diagram of an ultra-precision machine",
+        "texts": [
+            " The loop stiffness of the ultra-precision machine was predicted using the identified joint stiffnesses, and the validity of the identified joint stiffnesses was verified from the fact that the predicted loop stiffnesses well coincided well with the measured loop stiffnesses. This approach for identifying the joint stiffnesses of the hydrostatic guideways and bearings will be useful to more quantitatively evaluate the structural characteristics of the ultra-precision machine and to improve the performance of the hydrostatic guideways and bearings. The basic structure and design specifications of an ultraprecision machine for machining large-surface micro-features with side lengths of hundreds of mm and a depth of tens of \u00b5m are presented in Fig. 1 and Table 1. The ultra-precision machine has a gantry-type structure. The X-axis feed table moves in the transverse direction along the hydrostatic guideway mounted on the crossbeam connecting the upper ends of two columns. The Y-axis feed table moves in the longitudinal direction along the hydrostatic guideway mounted on the bed. The Z-axis feed table moves in the gravitational direction along the hydrostatic guideway mounted on the front side of the X-axis feed table. The C-axis rotary table rotates supported by the hydrostatics bearings mounted inside the Yaxis feed table",
+            " The eleven joint stiffnesses, which should be identified sequentially one by one based on the virtual prototype and the eleven measured compliances, are the normal and lateral stiffnesses of the X-axis, Y-axis and Z-axis hydrostatic guideways, the radial and thrust stiffnesses of the C-axis hydrostatic bearings, and the axial stiffnesses of the X-axis, Y-axis, and Z-axis linear motors. Since the Z-axis and C-axis feed systems are respectively mounted on the X-axis and Y-axis feed systems as shown in Fig. 1, the compliances of the Z-axis and C-axis feed systems are respectively affected by the joint stiffnesses of the X-axis and Y-axis feed systems besides the joint stiffnesses of the Z-axis and C-axis feed systems and the deformation of the structure elements. Therefore, in the joint stiffness identification process, the joint stiffnesses of the X-axis and Y-axis feed systems are firstly identified, and then the joint stiffnesses of the Z-axis and C-axis feed systems should be identified. In this study, to identify an interested joint stiffness Kj sequentially one by one, the optimization problem for minimizing the error e between the measured compliance Cm and the predicted compliance Cp was established as follows: 2 Minimize ( ) ( ) Subject to 0 j p j m j e K C K C K = \u2212 \u2265 (2) The predicted compliance Cp is the function of the joint stiffness Kj since it is estimated based on the virtual prototype expressed as Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003569_iros.2011.6048708-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003569_iros.2011.6048708-Figure3-1.png",
+        "caption": "Fig. 3. Validation of action-reaction principle. The deformations of lowfrequency object (green) and high-frequency object (red) are the same.",
+        "texts": [
+            " First, the validation of action-reaction principle has been performed using two beam models having the identical physical properties, each being calculated at different frequency. The low-rate curve-like object was fixed in the space, whereas the high-rate one was attached to the lowrate body on one end and to the haptic device on the other. Although each model was calculated at different frequency, the deformations of both objects was identical for arbitrary position of the haptic device as shown in Fig. 3. Second, we implemented a snap-in scene, depicted in Fig. 4. In this example, the high-rate object was represented by a stiff but deformable clamp attached to the haptic device via coupling spring. The clip was modeled with 20 corotational beam elements and nodes with positional and rotational degrees of freedom were used in the model. As the obstacle, a deformable cylinder composed of 103 tetrahedral elements was simulated. During the snap operation, up to 20 constraints were created. Various rigidities of both cylinder and clamp were tested with haptic feedback"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002514_jfm.2011.256-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002514_jfm.2011.256-Figure3-1.png",
+        "caption": "FIGURE 3. Geometry of a floating viscous sheet in a three-dimensional geometry.",
+        "texts": [],
+        "surrounding_texts": [
+            "3.1. Case of a floating, three-dimensional sheet with gravity, buoyancy and surface tension In the remainder of this paper, we apply our equations to a specific geometry: a thin, viscous sheet floating on a bath at hydrostatic equilibrium, with surface tension at the lower and upper interfaces. Each of the surface tensions \u03b3+ and \u03b3\u2212 at the upper and lower interfaces are assumed to be uniform. An in-plane flow of typical velocity U is imposed in the sheet, e.g. through in-plane loading. The density of the sheet material \u03c11 is smaller than that of the bath \u03c12. Let z = Hb be the height of the free surface of 126 G. Pfingstag, B. Audoly and A. Boudaoud the bath when the sheet is absent. The bath is an inviscid fluid at rest; the pressure in the bath is purely hydrostatic, p= \u03c12g(Hb \u2212 z), where g is the acceleration of gravity. The weight per unit volume of the sheet is simply (\u03c11g). Let z\u2217 be the typical vertical displacement of the interfaces, be it due to deformations of the initially planar mid-surface, or to variations of thickness. Surface forces applied at the interfaces come from hydrostatic pressure, of order (\u03c12gz\u2217), and capillary pressure, of order (\u03b3 z\u2217/L2). Here \u03b3 is the typical value of surface tension and L an characteristic in-plane length, such as the wrinkling wavelength, or a scale imposed by the in-plane flow. These surface forces can be compared to the typical viscous force per unit volume, (\u00b5U/L2), times the typical thickness z\u2217, using non-dimensional numbers. The Jeffreys number (Jeffreys 1925; O\u2019Keefe 1969) compares gravity and viscosity: Je= \u03c1gL2 \u00b5U , (3.1) where \u03c1 stands for the order of magnitude of any of the densities. Similarly, the inverse capillary number compares surface tension and viscosity: 1/Ca= \u03b3 \u00b5U . (3.2) A number of models could be considered based on these ingredients. We focus here on the case where gravity and surface tension both need to be considered in addition to viscosity. This is the most interesting situation since we have two factors that can limit the instabilities and which are potentially in competition: one relevant to large wavelengths and the other one to small wavelengths. Specifically, we consider two cases. (a) Both the Jeffreys Je and inverse capillary 1/Ca number are small; transverse forces should be considered of order \u03b5. The argument of \u00a7 2.5 suggests that we then consider short times and use the BNT scaling. The expressions of the applied forces read f (\u22121) z = 0, (3.3) f (1)z =\u2212\u03c11g+ \u03b4+(x, y, z)(\u03b3+)\u03ba+(x, y)+ \u03b4\u2212(x, y, z) \u00d7 [ (\u03b3\u2212)\u03ba\u2212(x, y)+ \u03c12g ( Hb \u2212 H + h 2 )] , (3.4) f (0)x = 0, (3.5) f (0)y = 0, (3.6) where \u03b4\u00b1(x, y, z)= \u03b4(z\u2212 z\u00b1(x, y)) denote the Dirac distributions centred on an interface z= z\u00b1(x, y), used here to account for surface forces, namely the capillary force and the force arising from pressure in the bath. In addition, we have introduced the following notation for the curvature of the lower and upper interfaces: \u03ba\u00b1(x, y)=\u22072z\u00b1(x, y)=\u22072 ( H(x, y)\u00b1 h(x, y) 2 ) . (3.7) We will refer to the case Je 1, 1/Ca 1 and t \u223c (z\u2217)2/(LU) as the BNT case. (b) Alternatively, we consider the case where both the Jeffreys Je and inverse capillary 1/Ca numbers are large; transverse forces should be considered of order 1/\u03b5, and are considered part of f (\u22121) z in our equations. Since the slope of the sheet is of order \u03b5, the projection of normal forces onto the (x, y) plane are of order \u03b50, and enter Linear and nonlinear stability of floating viscous sheets 127 into f (0)x and f (0)y : f (\u22121) z =\u2212\u03c11g+ \u03b4+(x, y, z)(\u03b3+)\u03ba+(x, y)+ \u03b4\u2212(x, y, z) \u00d7 [ (\u03b3\u2212)\u03ba\u2212(x, y)+ \u03c12g ( Hb \u2212 H + h 2 )] , (3.8) f (1)z = 0, (3.9) f (0)x = ( H + h 2 ) ,x (\u03b3+)\u03ba+(x, y)\u2212 ( H \u2212 h 2 ) ,x \u00d7 [ (\u03b3\u2212)\u03ba\u2212(x, y)+ \u03c12g ( Hb \u2212 H + h 2 )] , (3.10) f (0)y = ( H + h 2 ) ,y (\u03b3+)\u03ba+(x, y)\u2212 ( H \u2212 h 2 ) ,y \u00d7 [ (\u03b3\u2212)\u03ba\u2212(x, y)+ \u03c12g ( Hb \u2212 H + h 2 )] . (3.11) As explained in \u00a7 2.5, these large forces are studied at long times and using the Trouton scaling. We will simply refer to the case Je 1, 1/Ca 1 and t \u223c L/U as the Trouton case. These external forces are substituted into (2.44), (2.45) and (2.45). We keep notation concise using the number m= 0 for the Trouton case and m= 1 for the BNT case. We obtain the following set of equations. First, mass conservation remains h,t(x, y)+ (1\u2212 m) ( \u2202(u(x, y)h(x, y)) \u2202x + \u2202(v(x, y)h(x, y)) \u2202y ) = 0. (3.12) The in-plane force balance is obtained from (2.45), using the forces given in (3.11) and (3.6) integrated over thickness: Nxx,x + Nxy,y =\u2212(1\u2212 m) ( H + h 2 ) ,x (\u03b3+)\u03ba+(x, y)+ (1\u2212 m) ( H \u2212 h 2 ) ,x \u00d7 [ (\u03b3\u2212)\u03ba\u2212(x, y)+ \u03c12g ( Hb \u2212 H + h 2 )] , (3.13a) Nxy,x + Nyy,y =\u2212(1\u2212 m) ( H + h 2 ) ,y (\u03b3+)\u03ba+(x, y)+ (1\u2212 m) ( H \u2212 h 2 ) ,y \u00d7 [ (\u03b3\u2212)\u03ba\u2212(x, y)+ \u03c12g ( Hb \u2212 H + h 2 )] . (3.13b) Inserting the forces in (3.11) and (3.6) into (2.47), and carrying out evaluation of the slice-based operators, we find the membrane stress: Nxx(x, y)= 2\u00b5h[2(u,x + mH,xtH,x)+ v,y + mH,ytH,y] + (1\u2212 m)h \u03b3+ 2 \u03ba+(x, y)\u2212 (1\u2212 m)h \u03b3\u2212 2 \u03ba\u2212(x, y) \u2212 (1\u2212 m)h \u03c12g 2 ( Hb \u2212 H + h 2 ) , (3.14) 128 G. Pfingstag, B. Audoly and A. Boudaoud Nxy(x, y)= \u00b5h[u,y + mH,xtH,y + v,x + mH,ytH,x], (3.15) Nyy(x, y)= 2\u00b5h[u,x + mH,xtH,x + 2(v,y + mH,ytH,y)] + (1\u2212 m)h \u03b3+ 2 \u03ba+(x, y)\u2212 (1\u2212 m)h \u03b3\u2212 2 \u03ba\u2212(x, y) \u2212 (1\u2212 m)h \u03c12g 2 ( Hb \u2212 H + h 2 ) . (3.16) In the Trouton case, the out-of-plane force balance (2.46a) boils down to \u2212(\u03b3+)\u03ba+(x, y)\u2212 (\u03b3\u2212)\u03ba\u2212(x, y)\u2212 \u03c12g ( Hb \u2212 H + h 2 ) + \u03c11gh= 0 (m= 0), (3.17a) which is a partial differential equation for h and H. In the BNT case, we note that [K \u00b7 f (\u22121) z ](x, y) = 0: we substitute the remaining terms in the expressions (2.48) of the moment, namely M\u03b1\u03b2(x, y)=\u2212(\u00b5h3/3)(H,\u03b1\u03b2t + H,\u03b3 \u03b3 t\u03b4\u03b1\u03b2)/2, into the transverse balance (2.46b), which yields NxxH,xx + 2NxyH,xy + NyyH,yy = \u00b5h3 3 (H,xxxxt + 2H,xxyyt + H,y4t)\u2212 (\u03b3+)\u03ba+(x, y) \u2212 (\u03b3\u2212)\u03ba\u2212(x, y)\u2212 \u03c12g ( Hb \u2212 H + h 2 ) + \u03c11gh (m= 1). (3.17b) 3.2. Base state We consider a flat configuration, H(x, y, t) = H0(t) and h(x, y, t) = h0(t), which is spatially uniform but may evolve with time. The remainder of this paper is concerned with a local analysis of stability of this flat configuration. By \u2018local analysis\u2019, we mean that the typical scale of the in-plane flow is assumed to be much larger than the wavelength of the possibly unstable modes. The analysis is carried out over an intermediate length scale, where the strain rate D can be considered uniform. An in-plane flow can be present in this flat state, and can be driven by forcing at the boundaries. In the remainder of the paper, our main goal is to compute the growth rate of unstable modes, depending on the properties of this flow. To start with, we consider an in-plane flow that does not depend on time. This assumption can easily be relaxed: our expressions for the growth rate C as a function of the steady flow parameters \u03b1 and \u03b2 yield without modification the instantaneous growth rate as a function of the current flow parameters in the case of a time-dependent flow. This idea is implemented in full detail in \u00a7 5; see (5.2) in particular. Let us call \u03b1 and \u03b2 the principal values of the two-dimensional strain rate, ordered with the convention \u03b1 < \u03b2. (3.18) By convention, the x and y axes are aligned with the associated principal directions. This makes the strain-rate tensor diagonal, Dxx = \u03b1, Dxy = 0 and Dyy = \u03b2. With this choice of axes, the in-plane velocity is given by u0(x, y)= \u03b1x, (3.19) v0(x, y)= \u03b2y. (3.20) Linear and nonlinear stability of floating viscous sheets 129 On the right-hand side, we have omitted constants of integration representing a steady rigid-body motion, which do not affect the stability. Note that indices in parentheses, such as (0) in (2.32a) refer to the order in the expansion with respect to the small aspect ratio. By contrast, indices without parentheses refer to the expansion associated with the analysis of stability: in equation above, u0 refers to the base solution, and u1, introduced next, refers to the perturbation (marginally stable mode). Conservation of mass (3.12) reads, for our planar solution, h0,t(t)+ (1\u2212 m)(\u03b1 + \u03b2)h0(t)= 0. (3.21) In the Trouton case (m= 0), the second term is non-zero and the thickness h0(t) of the base flow must depend explicitly on time, except in the special case of a pure shear flow (\u03b1 + \u03b2 = 0). In the absence of interface curvature, the transverse equilibrium (3.17a) or (3.17b) imposes the following buoyancy condition: H0(t)= Hb \u2212 ( \u03c11 \u03c12 \u2212 1 2 ) h0(t). (3.22) Being homogeneous by assumption, the base state satisfies the in-plane force balance (3.13) automatically. The membrane stress is given by the constitutive law (3.16), which reads, after substitution of (3.22), N0 xx(t)= 2\u00b5(2\u03b1 + \u03b2)h0(t)\u2212 (1\u2212 m) \u03c11gh2 0(t) 2 , (3.23a) N0 xy(t)= 0, (3.23b) N0 yy(t)= 2\u00b5(\u03b1 + 2\u03b2)h0(t)\u2212 (1\u2212 m) \u03c11gh2 0(t) 2 . (3.23c) This unperturbed membrane stress will appear in the equations for the linear stability. 3.3. Perturbation We seek solutions to (3.12)\u2013(3.17) by perturbing this spatially homogeneous base state: h(x, y, t)= h0(t)+ h1(x, y, t), (3.24a) H(x, y, t)= H0(t)+ H1(x, y, t), (3.24b) u(x, y, t)= \u03b1x+ u1(x, y, t), (3.24c) v(x, y, t)= \u03b2y+ v1(x, y, t), (3.24d) where h0(t), H0(t), u0 = \u03b1x and v0 = \u03b2y characterize the flat base state of the previous section, and h1(x, y, t), H1(x, y, t), u1(x, y, t), v1(x, y, t) stand for the perturbations to thickness, mid-surface, and velocity. As an illustration, we consider the case where the in-plane velocity is imposed at the boundaries, which is far from the region of interest. In view of this, the perturbation to velocity should vanish at infinity: lim (x,y)\u2192\u221e u1(x, y)= 0, (3.25) lim (x,y)\u2192\u221e v1(x, y)= 0. (3.26) The stability is investigated next. In \u00a7 4 we treat the BNT case and find that the most unstable modes are purely undulatory (thickness h remains constant); in \u00a7 5 we study the Trouton case, and find that the unstable modes couple thickness and height variations. 130 G. Pfingstag, B. Audoly and A. Boudaoud"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001952_ichr.2009.5379580-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001952_ichr.2009.5379580-Figure8-1.png",
+        "caption": "Fig. 8. (a) The workspace projection of a simulated execution of a C-Space path. The TCP follows the path (white dots) by moving the hand toward the temporary goals (red dots) which are computed during execution. (b) A narrow view of an executed trajectory where workspace equivalents of C-Space path points (white), goals (red) and the executed trajectory (gray) are depicted. The displacement of the TCP, caused by an artificial lag, is quickly compensated.",
+        "texts": [
+            " In Fig. 7 the search for a new goal starts at C~obot and both goal limits gc and gw are calculated. Although Vc would allow robot to move until gc, this would violate the workspace velocity limit V w which means that there would be a joint moving too fast. So g w will become the next temporary goal gnext. The direction of further movement in C-Space is given by v = gnext - Cro bot , and in case v does not violate C-Space or workspace speed limits, it can be passed to the low level controllers. In Fig. 8 the results of two simulation experiments are shown. The left figure shows the workspace TCP positions during the execution of the C-Space path. The white dots and the black line depict the path of the TCP in workspace and the red dots are placed at the workspace positions of the temporary goals which were computed during execution. The right figure shows the result of an experiment where an artificial lag of 1000 ms forces the high-level control loop to stop its execution. Since the low-level controller is still executing the last calculated velocity values, the TCP moves away from the trajectory"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003776_j.mechmachtheory.2012.03.003-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003776_j.mechmachtheory.2012.03.003-Figure3-1.png",
+        "caption": "Fig. 3. Four known solutions of the Bennett linkage with respect to angles \u03c71 and \u03c72.",
+        "texts": [
+            " (9), and the second equation was obtained after substituting \u03c61=\u03c62=\u03c0 into the same system. Among these four possible solutions, condition (10) is fulfilled by two similar equations: \u03b61 \u00bc \u03c72 and \u03b62 \u00bc \u03c71; \u00f011\u00de both obtained on the basis of U1z=U2z and U1z=\u2212U2z. If we connect the tori radii with revolute pairs at points Q, R1, S1=S2 and R2, then, on the basis of the system of Eq. (7), we obtain a four-bar spatial Bennett linkage. There are four known solutions for this mechanism with respect to angles \u03c71 and \u03c72 [10]: \u2212 \u03c71>0 and \u03c72>0 or a symmetrical solution \u03c71b0 and \u03c72b0 (Fig. 3A, B); \u2212 \u03c71>0 and \u03c72b0 or a symmetrical solution \u03c71b0 and \u03c72>0; (Fig. 3C, D); They result from the solution for the system of Eq. (7) and are equivalent in pairs. The function of the position of symmetric solutions (Fig. 3A, B, C, and D) remains identical. The solutions for the system of Eq. (10) also remain identical (16) for both variants of the mechanism, which are both obtained on the basis of condition \u03c71\u03c72>0 and \u03c71\u03c72b0. There are also two solutions for the system of Eq. (8):r2=r1, s2=s1, \u03c72=\u03c71 \u2014 trivial identity solutions,and r2 \u00bc r1; s2 \u00bc s1; \u03c72 \u00bc \u2212\u03c71: \u00f012\u00de The system of Eq. (10) does not have any solutions to satisfy Eq. (12). The Myard mechanism is obtained [12] after applying the universal joint in point S"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002348_1350650111427508-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002348_1350650111427508-Figure4-1.png",
+        "caption": "Figure 4. Aerodynamic bearing mounted in the housing.",
+        "texts": [
+            " In addition, the rotor has a large tilting stiffness because the two Lomakin bearings are largely spaced. The angled injection ensures a very low or negative cross-coupling stiffness, leading to a large effective damping. Hence, the test rig was operated up to 60 kr/min (the largest speed enabled by the spindle) without any sign of instability. The rotor is considered rigid because the first natural frequency of its freely supported mode is largely higher than the maximum attainable rotation speed. The tested bearing is made of graphite serrated in a stainless steel sleeve (Figure 4). Two conical parts are used for mounting the cylindrical sleeve in the housing. The static and the impact forces are applied on this housing that is freely floating on the rotor. In order to avoid the dynamic misalignment problem exhibited during previous tests,14 the housing is mounted on three flexible stingers fixed on a base plate. The stingers do not interfere with measurements because their stiffnesses are one order of magnitude lower than those of the tested bearing. The static load is vertically applied on the bearing by a manually driven bolt and nut system and spring"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001417_s00791-008-0109-x-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001417_s00791-008-0109-x-Figure10-1.png",
+        "caption": "Fig. 10 Formation of insertions and swallow tails in the process of front development",
+        "texts": [
+            " The value function near the point d outside the set M is continuous. In the algorithm, the boundary \u2202W(\u03c4i , M) of the set W(\u03c4i , M) is represented by a finite collection of points connected by straight segments. The construction of the current front F(\u03c4i+1) is based on the consideration of extremal trajectories related to the previous front and on the application of procedure for the removal of \u201cswallow tails\u201d. Let us explain the removal procedure. Let F1 be a polygonal line in the plane, the positive and negative sides of the line being marked (Fig. 10). The control in the system strives to push the polygonal line in reverse time in the direction of the negative side. We order the vertices of the original polygonal line F1 in the clockwise direction and consider outward normals to the line links. The trajectories of the controlled system, which are extremal with respect to these normals will be emitted from points of F1 on a small reverse-time interval. Since some points generate more than one extremal trajectory, the ends of the trajectories can give a polygonal line with local self-intersections. In Fig. 10, each of the vertices a, b, c, and e of the initial polygonal line generate two extremal trajectories. The vertices a and c of the line F1 are the points of local convexity. Therefore, the emitted trajectories form an additional segment (an insertion) on the new line. The vertices b and e are the points of local concavity. In this case, we obtain two self-intersections of the line composed of the ends of extremal trajectories. Such self-intersections are used to be called \u201cswallow tails\u201d. Since they are located in the positive region with respect to the advancing line that we are looking for, they should be removed. The obtained polygonal line is the thick solid line F2 in Fig. 10. Various versions of algorithms for the construction of solvability sets of time-optimal problems and problems with fixed terminal time that use procedures for removing swallow tails were developed by the authors beginning with the paper [25]. The papers [3,10] are to be mentioned as earlier works which utilized the concept of removing the swallow tails for the numerical construction of solvability sets. To maximally simplify the logic of the swallow tail removing in the algorithm considered, we separate the procedure for computing the extremal trajectories into two stages by accounting for the specifics of the right-hand side of system (11)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003710_gt2012-68956-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003710_gt2012-68956-Figure4-1.png",
+        "caption": "FIGURE 4. Test arrangements (A: blow down testing; B: Axial width testing; C: Stiffness testing)",
+        "texts": [
+            " The mass flow can be metered for each test setup. The test facility is capable of conducting three different types of tests. First, the blow down behavior of brush seals can be observed. It is well known that the bristles are bent in radial direction by aerodynamic forces, closing the initial gap (as-built clearance) between the bristle tips and the rotor surface [2-5]. For testing purposes, the distance between the bristle tips and the backing plate can be monitored with a digital microscope (see FIGURE 4 A), while the pressure difference can be adjusted up to 6 bar. The second type of tests considers the axial widths of the bristle pack, which is influenced by the pressure difference across the brush seal. As can be seen in FIGURE 4 B, the digital microscope is used for monitoring the brush again. Here, a mirror is used to look in radial direction. Third, the test rig is capable of measuring the stiffness of the bristle pack (see FIG- URE 4 C). Therefore, a sensing device, which is connected to a load cell, can be moved into the bristle pack while monitoring both force and deflection. Contrary to Bidkar et al. [6], the test specimen of the Braunschweig test rig is a finite-extent shoe. It is well known that this testing setup leads to results containing endeffects",
+            " Bristle bending effect 0,0 1,0 2,0 3,0 4,0 5,0 6,0 7,0 8,0 1 2 3 4 5 6 condition 1 condition 2 condition 3 condition 4 no n di m en si on al m as s flo w m /m 0 \u2206p in bar m0 FIGURE 13. Leakage flow distribution of a straight bristle pack under different pressure drops the bristle pack halves. The clearance distribution (see FIGURE 10) is metered at the downstream bristles while the behavior of the upstream bristles cannot be analyzed within this test setup. Therefore, the axial width setup, (see FIGURE 4 B) was used to find out how the upstream bristles may affect the leakage flow distribution. It was found that the upstream bristle pack halves behave as separate brush seals. While the downstream bristle pack halves show the blow down behavior depicted in FIGURE 10, the upstream bristles are able to show a blow down and therewith close the gap between the bristle tips and the rotor disc. Thereby, the clamping forces, which influence the gapping of the bristle pack, also influence the amount of the blow down, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001408_robot.2009.5152309-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001408_robot.2009.5152309-Figure2-1.png",
+        "caption": "Fig. 2. Tip over distances of different configurations",
+        "texts": [
+            " In this condition, any disturbance will result in a restoration force, which tends to return the body to the equilibrium condition [7]. In mobile robots, the first concern about stability is to avoid the system to tip over. 978-1-4244-2789-5/09/$25.00 \u00a92009 IEEE 1340 1) Stability Margin: In early works of walking machines [15], the vehicle is considered stable if the CM projection lies inside the support polygon. The support polygon (SP) is obtained from the tip over axis ati \u2208 R3 which connects all m terrain contact points. The distances between the projections of CM and SP (see Fig. 2) are referred as the tip over distances dti . The minimum tip over distance is defined as the robot stability margin: \u03b1 = min(dti), i = 1, ...,m (2) According to this criterion, the configuration with maximum stability is achieved when all dti have the same length. The stability margin does not explicitly consider the height of the robot CM. Two robots with same dti have the same stability margin, independent of ground clearance. 2) Gradient Stability Margin: This criterion1, proposed in [17], was developed for mobile machines with manipulators used in construction, mining and forestry"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001470_j.jsv.2008.05.001-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001470_j.jsv.2008.05.001-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of the double pendulum with a stopper at the lower hinge.",
+        "texts": [
+            " The decay rate of energy and angular momentum show dependencies unique to the case on hand. r 2008 Elsevier Ltd. All rights reserved. It has been established that torque proportional to the angle between the shank and the vertical stabilises a biped mechanism [1]. A self-impacting double pendulum is a structural component of the gait cycle of a biped with the impact being at the knee and has been studied here. The response of a double pendulum without impacts is chaotic in nature [2]. A stopper at the lower hinge mimics the knee joint impact of a leg, as shown in Fig. 1. This self-impact may or may not curb the chaotic behaviour (intrinsic to the nonlinearity in the system), depending on the system parameters, the initial conditions and the nature of impact (elastic or otherwise). An impact model has been used to analyse the response of a self-impacting double pendulum with varying coefficient of restitution. A methodology of modelling impact with energy losses amenable to the time domain analysis has been proposed. The method, limited only by the computational ability of the solver, assumes significance as the accuracy of solution of the nonlinear system resulting from the impact is driven by the accuracy of the impact modelling"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001882_j.enconman.2009.04.029-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001882_j.enconman.2009.04.029-Figure1-1.png",
+        "caption": "Fig. 1. The equivalent current space phasor.",
+        "texts": [
+            " Nomenclature Ts period of time that every voltage vector is applied v0 to v5 voltage vectors applied by the inverter is0 to is5 the stator current space vectors corresponding to v0 to v5 hr the actual rotor position ~hr the estimated rotor position wm the flux linkage produced by the permanent magnet w\u0302m magnitude of wm ws the flux linkage space phasor of the stator is the stator current space phasor isq the quadrature axis stator current isd the direct axis stator current vq the quadrature axis stator voltage vd the direct axis stator voltage ise the equivalent current space vector i\u0302se magnitude of ise ide the direct axis component of the ise iqe the quadrature axis component of the ise Ls the inductance matrix of the stator Ld direct axis component of Ls Lq quadrature axis component of Ls rs the stator resistance xr speed of the rotor The relations of the stator flux linkage in the d\u2013q reference frame are wd \u00bc Ldid \u00fe w\u0302m wq \u00bc Lqiq \u00f02\u00de the magnet\u2019s contribution to the stator flux linkage can be considered as an equivalent current along the d-axis as follows: w\u0302m \u00bc Ldi\u0302m \u00f03\u00de substituting (3) in (2), an equivalent current space vector ise can be introduced, as shown in Fig. 1, causing whole of the stator flux linkage which it\u2019s d-q components as follows: ide \u00bc id \u00fe i\u0302m iqe \u00bc iq \u00f04\u00de The saturation level is a function of the magnitude of ise i\u0302se \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi i2 de \u00fe i2 qe r \u00f05\u00de So the d\u2013q components of the stator inductance are the functions of i\u0302se as shown in Fig. 2, therefore, the d\u2013q components of the stator flux linkage can be written as wd \u00bc Ld \u00f0\u0302ise\u00de ide wq \u00bc Lq \u00f0\u0302ise\u00de iqe \u00f06\u00de As shown in Fig. 1, the angle between stator current phasor is and the axis of the magnet affects the magnitude of the equivalent current phasor i\u0302se, and the saturation level, consequently. Assuming the initial rotor position is hr , a voltage space vector Vs\u00f0h\u00de is applied to the motor, and the consequent current space vector is is as shown in Fig. 1. From (4) and Fig. 1, if q decreases, the ide increases, consequently the saturation level increases, therefore Ld decreases that it causes to increase is. On the other hand, if the angle of the applied voltage vector Vs\u00f0h\u00de closes to the north pole (h! hr ; Fig. 1), then the corresponded current vector is becomes larger. So by applying several voltage vectors to the motor and measuring the consequent currents, the rotor position can be estimated as the angle of the voltage vector which its corresponded current is maximum. By using this principle, the following method is proposed to detect the initial rotor position: 1. In the first step, six voltage space vectors v0(0 ), v1(60 ), v2(120 ), v3(180 ), v4(240 ), and v5(300 ) are applied to the motor as shown in (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002285_660432-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002285_660432-Figure1-1.png",
+        "caption": "Fig. 1 - Sketch of basic mechanism for generating relative",
+        "texts": [
+            " However, since the primary interest here is self-lock - ing bolts, it must be reported that no success has been obtained with the process of fluctuating tension, which so often is pointed out in theoretical machine design as the cause of bolt loosening. While it is true that, in many machine situations, fluctu- ating tension remains the primary component of load on a bolt, there are also many situations in which two mating parts held by a bolt rotate a minute amount with respect to one another. Under such conditions the bolt may be observed to loosen and eventually come out. This can be demonstrated in the laboratory in at least two specific ways, both of which utilize the basic phenomena illustrated in Fig. 1. Here, one may see that two primary elements are needed in order to cause loosening of a bolt by oscillating torque or rotation. The first of these is the stationary portion, marked A in Fig. 1, composed of the base rigidly attached to some massive foundation. To this base is attached an ear E having a body hole so that the bolt B may be put through the hole and screwed tightly into the part C. This simulates the attachment of one part to another by a bolt, in the usual way. At this point two possible mechanisms of loosening are available. In the first, described in Ref. 1, a small fluctuating angular displacement is imposed on the part C, generally of harmonic form A e sin wt, and the resulting motion of bolt B is observed",
+            " To the best of the writer's knowledge, this phenomenon was first described in Ref. 1. DESCRIPTION OF EQUIPMENT Since a typical machine element in which a self-locking bolt is used is not generally subjected to tension conditions alone, it seems quite possible that small relative rotations between parts may in many cases be responsible for the loosening of bolted connections. In order to define clearly the effects of relative rotation, a machine was designed and constructed similar in principle to that shown in Fig. 1, but with some added instrumentation and measuring devices. A photograph of this machine is shown in Fig. 3; Figs. 4 and 5 show detailed views of the bolt B, ear E, and head section C of the machine. Fig. 5 also illustrates the measuring device used to indicate looseness of the bolt. In the design of this machine, it should be noted that an internal strain gage dynamometer is built into the ear. This dynamometer is used for measuring axial load in the bolt as it is tightened into the head section C",
+            " However, the large forces carried here cause elastic deformations, which are of the same order of magnitude as the preset eccentricities when these eccentricities are small, so that in this case the actual values of D e are very difficult to measure. Under such conditions, considerable doubt exists as to the validity of the data in this region, and primarily for this reason one finds difficulty in fully accepting the results of tests from this machine at low eccentricity values. In order to avoid this difficulty of unknown total eccentricity, it was decided to construct a machine operating on principles similar to that shown in Fig. 1, but now modified in such a way that a fluctuating torque would be applied to the member C of Fig. 1. If such a fluctuating torque could be applied independently of the angle of rotation of part C, then all doubt would be removed concerning the exact variables used in a particular test, The fluctuating torque chosen at the beginning of the test should remain constant throughout the duration of the test, independent of the magnitude of the resulting oscillatory motion. In order to accomplish this, it was necessary to utilize some sort of device that was capable of exerting forces independent of oscillatory amplitude",
+            " As a crack progresses in a typical fatigue specimen and amplitude of displacement increases, the forces will indeed remain constant until fracture is complete. Such machines are typically equipped with a microswitch displacement shutoff control so that at large displacements, coinciding with the fracture of a specimen, the machine automatically shuts off. In this particular instance, a 100 lb capacity Baldwin -Sonn- tag fatigue testing machine was equipped with its standard torsion fixtures so modified as to produce the type of action illustrated in Fig. 1, except that now part C of Fig. 1 was subjected to an oscillatory torque AT sin cot rather than an oscillatory displacement. A photograph of the modified fixture is shown in Fig. 6, and the torque or moment arm is shown in the background of Fig. 7. From these photographs it may be seen that the small platen which moves up and down in an oscillatory fashion, transmitting the unbalanced eccentric force, is coupled by means of a torque arm directly through a part similar to arm C of Fig. 1. This, in turn, is held to a flat plate by means of the test bolt. The flat plate is securely fixed to the base of the testing machine by means of two heavy supports, as shown in Fig. 7. Fig. 8 shows an element of the test fixture corresponding to part C of Fig. 1. This rod is, of course, threaded to receive the test bolt and is further specially constructed to measure the axial load in that bolt. The form of this construction is shown in Fig. 9, where it is seen that a portion of the cylindrical member is removed so that a thin cylindrical shell can be substituted. This shell has been instrumented with strain gages so that axial load may be accu- rately measured. TESTING PROCEDURE The procedure for testing a bolt under a particular set of conditions begins by inserting the bolt through the fixed mem - ber and into the threaded cylindrical portion, followed by tightening the bolt to some level of shank stress as indicated by the strain gage instrumented dynamometer on the shoulder of the cylindrical portion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000336_iecon.2006.347699-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000336_iecon.2006.347699-Figure1-1.png",
+        "caption": "Fig. 1 Outline of one segment of the Linear Synchronous Motor",
+        "texts": [
+            "00 '2006 IEEE This paper is organized as follows. In Section II the linear synchronous motor is described and modelled. In Section III the proposed sensorless method is presented. In Section IV experimental results are shown. Finally, in Section V, some conclusions are drawn. The Long-Stator Linear Synchronous-Motor under consideration is arranged in multiple segments i.e. stators. Each segment is composed of two facing stator-sides forming a slot between them, where the mover translates. A drawing of one segment is shown in Fig. 1. The windings of both sides are parallel-connected and electrically independent from the other segments. One inverter is therefore used to drive each segment. Though several segments compose the carriageway, only two must be fed for each vehicle: the segment where the mover currently is, and the next one nearer to the mover. As can be appreciated in Fig. 1, the mover has surface magnets, so the magnetic saliencies are mainly due to the saturation they produce in the stator. Each segment of the linear synchronous motor can be modelled in the stationary reference frame, while the mover is in this segment, as follows: ( ) [ ]sin cos TPMdR v dt \u03c0\u03bb= + + \u2212 \u03b8 \u03b8 \u03c4 u i L i (1) where: x\u03b8 = \u03c0 \u03c4 (2) and [ ]Tu u\u03b1 \u03b2=u is the voltage vector, [ ]Ti i\u03b1 \u03b2=i is the current vector. Variables x and v are the mover\u2019s position and speed, respectively. The resistance is designated by R , the inductance matrix by L , the flux linkage due to the permanent magnets by PM\u03bb , and the pole-pitch by \u03c4 ",
+            " in the middle of Segment 2, Controller 1 can be detached from Segment 1 and attached to a third segment into which the mover will enter. This procedure will not affect the position detection because the detached segment was not providing position information at that time. IV. EXPERIMENTAL RESULTS An experimental setup was used to validate the proposed sensorless method. It is composed of a linear motor, two inverters, two PWM and ADC interface cards, and a controller implemented with an x86 PC. The linear PM synchronous motor is arranged in two segments, where each one is as described in Fig. 1. The most significant parameters of the motor and the setup are presented in Table I. The actual and observed position of the mover is shown in Fig. 6. The starting point is in Segment 1 and at 320x mm= the mover passes over the segment transition. No important deviation of the observed position can be appreciated on the transition. The higher deviation happens when accelerating or decelerating, as shown in the zoomed section. It should be noted that the only load on the mover is due to the cogging force and the acceleration of the mover\u2019s mass"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003352_9781118067147-Figure6.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003352_9781118067147-Figure6.1-1.png",
+        "caption": "Figure 6.1 The surface plasmon resonance principle in a typical setup. Reproduced with permission from [4].",
+        "texts": [
+            " The principle of SPR, known since 1950s [1\u20133], was not commercialized for biological sensing purposes until the late 1980s when the first BiacoreTM instrument was launched on the market [4]. This had been preceded by scientific investigations in which the design of surface-sensitive interactions on a metal film grafted on an optical prism for SPR detection was done [5]. Biomechatronic Design in Biotechnology: A Methodology for Development of Biotechnological Products, First Edition. Carl-Fredrik Mandenius and Mats Bj\u20acorkman. 2011 John Wiley & Sons, Inc. Published 2011 by John Wiley & Sons, Inc. 85 The optical principle of SPR is displayed in Figure 6.1. An optical beam, at a selected wavelength, impinges on a quartz prismwhere it reflects on ametal film covering the prism\u2019s upper surface. At a certain incident angle specific for the material, the energy propagates in parallel with the film. The energy corresponds to the quantum required for excitation of the metal\u2019s p-electrons (the so-called surface plasmon state). This quantum depends on the properties of the metal and refraction index of the surrounding media. Biomolecules covering thefilm influence the refractive index value"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003738_iros.2011.6094426-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003738_iros.2011.6094426-Figure1-1.png",
+        "caption": "Fig. 1. (a) The fiber directions of the two fiber-reinforced composite (FRC) layers are oriented at angles of \u00b1, to the longitudinal axis of the actuator. This creates anisotropy in the compliance matrix and allows exploitation of extension-twisting coupling to create a twisting actuator. (b) A three-dimensional and end-on view of a twisting actuator with length L, width W, and twist angle (}twist. (c) A cross-sectional view of an actuator, showing piezo layer thickness tp, composite layer thickness te, the z-axis defined with z = 0 at the midplane, and electrical connections to the top and bottom composite layers.",
+        "texts": [
+            " The authors are with the School of Engineering and A pplied Sciences and the Wyss Institute for Biologically Inspired Engineering, Harvard University, Cambridge, MA 02138. bfinio@fas . harvard. edu 978-1-61284-456-5/11/$26.00 \u00a92011 IEEE 384 Such actuators could be used for a variety of applications from precision micro manipulation to high bandwidth power delivery. Twisting motion is achieved with a single piezoelectric layer by laminating anti symmetric top and bottom fiber reinforced composite (FRC) layers (Fig. la). Thus an actua tor of length L and width W can achieve a rotation angle of Btwist (Fig. 1 b) by applying a voltage across the piezoelectric layer (for d31 mode actuation), using the conductive FRC layers as electrodes (Fig. lc). The following section presents derivation of the theory that predicts output twist angle, blocked torque and energy density based on the geometry and arrangements defined in Fig. 1. The model for a clamped-free piezoelectric cantilever actu ator consisting of an arbitrary number of active piezoelectric and passive composite layers in arbitrary orientations is originally presented in [3], derived from the information in [11]. Here we use the same approach to derive the constitutive equations for unimorph twisting actuators. First, assuming a state of plane stress for each layer (i.e. stresses act only in the x -y plane since there are no surface tractions), the in-plane stresses are given by where the subscripts 1, 2 and 3 correspond to a local coordinate system oriented with the fiber direction (Fig",
+            " For an anisotropic material (the FRe), elements of [Q] are defined as Q11 = El (5) 1 - V12V21 Q12 = V12E2 (6) 1 - V12V21 Q22 = E2 (7) 1 - V12V21 Q66 = G12 (8) where El is the Young's modulus in the fiber direction, E2 the modulus in the cross-fiber direction, and the following relationship holds: (9) To account for arbitrary orientation of anisotropic material layers, the stiffness matrix is modified: Q16] ([ Ex ] [d31] ) 2:: ;: y d\ufffd2 E3 (10) where [Q ij] is the adjusted stiffness matrix with the property (11) with the transformation matrix [T] defined as (12) where m = cosh) and n = sinh) (Fig la). To calculate overall deflections (displacements and/or ro tations of the output), midplane strains and curvatures are integrated over the appropriate dimension. These strains and curvatures are given by where EO = [Ex Ey Exyf, and K = [Kx Ky Kxy]T. A, Band D are elements of the stiffness matrix defined as (14) n (15) n (16) where Zn is the height of the nth laminate layer with respect to the midplane of the structure (Fig. 1 b). Next and Mext are the external forces and moments per unit width (i.e. N [Nx Ny Nxy]T, M [Mx My Mxyf), and NP and MP are the piezoelectric forces and moments per unit width, defined as: [Ni(E3)JP = \ufffd 1\ufffd\ufffd1 [Qij]nd3jE3dz (17) [Mi(E3)]P = \ufffd 1\ufffd\ufffd1 [Qij]nd3jE3zdz (18) Given Eq. 13, the maximum angle of twist, blocked torque, and energy density can be calculated for a given actuator geometry. T he curvature element Kxy is related to the vertical displacement of the actuator Z as Kxy = ;/;;y' therefore the angle of twist is calculated by evaluating the slope \ufffd\ufffd at the end of the beam: \ufffd1 (r5Z1 ) \u00b0twist = tan r5y x=L (19) where L is the actuator length",
+            " Laser cut-files are made with a 2D CAD program (DWGEditor, Dassault Systems) and the laser system is operated with ProLase XP (Amer ican Laserware Inc.). Rectangular pieces of desired length, width, and angle in the case of the composite layer, are cut from bulk piezoelectric and carbon fiber composite sheets available in discrete thicknesses from the manufacturer (PZT5H from Piezo Systems Inc. and M60J from Toray America respectively). Two carbon fiber layers and one piezoelectric layer are stacked to form the layup shown in Fig. 1 and vacuum bagged. The carbon fiber is pre-impregnated with a heat-curing epoxy, and the structure is cured under heat and pressure (with a Cascade TEK TFO-l oven), bonding the layers together. Actuator blocked torques (torque at zero rotation) were measured in a clamped-clamped configuration with a 6-axis Nano 17 force/torque sensor (ATI Industrial Automation). Data was recorded with data acquisition (DAQ) hardware and software supplied by ATI with the Nanol7. Custom mechanical mounts were used to connect the actuators to mechanical ground and to interface with the sensor, and an XYZ micrometer stage and Labjack (Thorlabs Inc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure11-1.png",
+        "caption": "Fig. 11 Contact pressure of meshing teeth",
+        "texts": [
+            " Moreover, with the linear distribution in modified gears, especially along single and double meshing zones, the instantaneous changes in the F/b ratio can be eliminated from the damaging effects. Based on the fact that the value of loading in single meshing zone is twice that at the double mesh, the twofold increase in gear tooth width, b, in the single meshing zone will cause a corresponding decrease in contact pressure at the rate of around 40 to 50% as calculated from (1). That is, the contact pressure values between points DB will fall by 40 to 50%. Also, the shaded area in the contact pressure diagram (Fig. 11) will disappear. 3.5 Temperature distribution along the contact path Metallic gears generally operate in oily media, whereas plastic gears run dry. In this respect, metallic gears have the advantage of dissipating heat out of the gear teeth. Plastic gears have no such advantage, and due to their lower thermal conductivities, the gears are subject to damages as a result of tooth temperatures. A large portion of the heat produced in gear teeth is due to friction and the transmitted load per unit tooth width is directly proportional to the friction-based heat produced"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000667_s00170-007-0927-x-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000667_s00170-007-0927-x-Figure2-1.png",
+        "caption": "Fig. 2 Sectional view of the friction force measurement setup",
+        "texts": [
+            "0838 Dynamic viscosity of base oil with concentration ratio at 5% additive \u03b7 (Pa at 20\u00b0C) 0.0862 Appearance color of oil fortifier (ASTM D1500) Clear, light amber Viscosity at 40\u2013100\u00b0C of oil fortifier (ASTM D2270) 70\u20139 cst Density of oil fortifier (ASTM D942) 1.05 g/ml Acid number of oil fortifier (ASTM D 2896) 0.4 mgKOH/g power stroke load transmitting through the connecting rod is more than the other strokes and inertia force created by various parts of an engine like pistons, connecting rods, flywheels and crankshafts. Figure 2 shows the details of the construction and working principle in a test rig. The journal is assembled with a bearing plate (measurement plate-5), which consists of oil supplier-3 to the contact surfaces. The measurement plate is connected with a pull lever with a spherical head-06 at two points which are 180\u00b0 in opposite direction. There are two transducers-02 arranged in the other direction of the journal on the horizontal line passing through the centre. The other ends of the lever of pull-06 are connected with measurement beam-09 at the fulcrum points"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003491_j.engfailanal.2013.09.019-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003491_j.engfailanal.2013.09.019-Figure4-1.png",
+        "caption": "Fig. 4. Lubricated disks in contact (co-ordinate system).",
+        "texts": [
+            " In ideal case it can be assumed that the gear surfaces are separated by a lubricant film generated by the gear rotation. The gear rotation causes pumping of lubricant flowing around the gears in the direction of rotation. Furthermore, as reported in standard tribology text books, two mating gears in contact can be represented by simplified geometry of two rollers as shown in Fig. 3 [15]. Thus, the theoretical elastohydrodynamic lubrication (EHL) model is developed for two lubricated disks in contact as shown in Fig. 4. The lubricant pressure distribution as a function of speed, contact geometry and lubricant viscosity and density can be described by one dimensional Reynolds equation: @ @x qh3 12 g @p @x ( ) \u00bc @ @x qh u1 \u00fe u2 2 n o ; \u00f01\u00de where x e (xl, xc) and u1\u00feu2 2 \u00bc us. Using the following dimensionless quantities: X \u00bc x b ; P \u00bc p PH ; H \u00bc h b ; g \u00bc g g0 ; q \u00bc q q0 ; H0 \u00bc h0 b ; U \u00bc g0us ER ; W \u00bc w ER ; G \u00bc aE: Eq. (1) reduces to the non-dimensional form: @ @x qH3 g @p @x ( ) \u00bc k @ @x fqHg where; k \u00bc 12g0usR 2 b3pH : \u00f02\u00de The boundary conditions for Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000762_tsmcb.2007.909943-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000762_tsmcb.2007.909943-Figure3-1.png",
+        "caption": "Fig. 3. Capillary force parameters during a sphere and flat surface contact.",
+        "texts": [
+            " It has been shown that increasing the surface roughness decreases the van der Waals forces [4]. Thus, taking the surface roughness into consideration, as shown in Fig. 2, the van der Waals force is expressed as [6] Fvdwb = ( z z + b/2 )2 Fvdw (5) where z is the distance, b is the height of the surface irregularities, and Fvdw is the van der Waals forces between the plane plate and the sphere. In ambient operational environment, the water layer is present on the surface of the sphere and the substrate. A liquid bridge occurs between them at close contact, as shown in Fig. 3. In [20], the macroscopic theory of capillarity is proven to be applicable for a curvature radius on the order of molecular size. Assuming that 1) r p Rp; 2) the surfaces are coated with a film of constant thickness e; 3) the contact angle is 0, which should be the true in our case; and 4) the surface attraction through the liquid phase is negligible, the capillary force can be written as [21] F cap = 4\u03c0\u03b3Rp ( 1 \u2212 h\u2212 2e 2r ) (6) where \u03b3 is the liquid (water) surface energy, e is the thickness of the water layer, and r is the radius of curvature of the meniscus, as shown in Fig. 3. Moreover, the volume of liquid condensed in the bridge and the film thickness distribution can also influence the capillary force, but it can be ignored in our case [21]. The capillary forces for the bp and bs interactions can be calculated from (6). It is important to notice that baking the sample before the manipulation process can greatly reduce the capillary forces [21]. For the electrostatic force, Coulomb forces are considered only. Using the point charge Assumption, the electrostatic force between an uncharged metal wall and a charged sphere is given by F elec = \u03b50\u03c0d 2 ( 3\u03b51 \u03b51 + 2 )2 E2 (7) where \u03b50 and \u03b51 are the dielectric constants of free space and the material, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000017_6.2005-6002-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000017_6.2005-6002-Figure5-1.png",
+        "caption": "Figure 5. Front view of the 2-frame and body frame",
+        "texts": [
+            " ( )[ ] comVcom FLf rr \u03b8\u03c8 ,22 = (42) American Institute of Aeronautics and Astronautics 12 (43) ( )[ ] \u23a5 \u23a5 \u23a5 \u23a6\u23a2 \u23a2 \u23a2 \u23a3 \u2212 \u23a5 \u23a5 \u23a5 \u23a6\u23a2 \u23a2 \u23a2 \u23a3 = 1 0 0 0 cos sin cos 0 sin 0 1 0 ,2 \u03c8\u03c8 \u03b8\u03b8 \u03b8\u03c8VL \u23a4\u23a1\u23a4\u23a1 \u2212 0 sin cossin 0 cos \u03c8\u03c8\u03b8\u03b8 ( )[ ] comcomBcomB fLf rr 22 \u03c6= (44) The components of the commanded vectors in eq. (42) and (44) are given by , \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 = com com com z y x com f f f f 2 2 2 2 r \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 = com com com zB yB xB comB f f f f r The situation after rotation into the 2-frame is shown in figure 5. Figure 5 depicts the view from the front of the aircraft, with the X-axis coming out of the page. From figure 5, the bank angle command is constructed as ( ) comcom zycom ffa 22 ,2tan \u2212=\u03c6 (45) From figure 5, the body frame z-axis specific force command is given as 2 2 2 2 comcomcom zyzB fff +\u2212= (46) The body frame x-axis specific force command is (47) comcom xxB ff 2= IV. Adaptive Autopilot We design adaptive controllers for tracking the normal acceleration command , lateral acceleration and bank angle command comzB f 0= comyB f com\u03c6 . The throttle controller is a PI controller with an anti-windup feature13 for tracking speed command V or a command formed by a combination of the longitudinal acceleration com American Institute of Aeronautics and Astronautics 13 command and "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000300_iros.2006.282182-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000300_iros.2006.282182-Figure4-1.png",
+        "caption": "Fig. 4. Hybrid control approach.",
+        "texts": [
+            " That is, a motion such that the leg would rotate in the hip while maintaining its fully stretched configuration. Note that such motion is typical for human walk [4]\u2013[6]. Fortunately, there is a straightforward solution to this problem. Whenever stretched-leg motion is needed, ankle joint control can be directly applied. In this way, we obtain a hybrid control framework, such that approach and departure from the singularity is done via the SingularityConsistent method, while motion along the singularity (outer boundary) is controlled directly in joint space (see Fig. 4). Here we describe the implementation of our hybrid control approach for achieving static walk. We focus thereby on two main points: (1) realization of the hybrid control approach by means of motion primitives, and (2) the importance of static stability margin. Statically stable cyclic walk will be generated by repeating a fixed cycle of motion primitives. These motion primitives are defined as follows (see also Fig. 5): 1) Initial position. 2) Move CoM over the left foot. 3) Lift the right foot"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003352_9781118067147-Figure6.7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003352_9781118067147-Figure6.7-1.png",
+        "caption": "Figure 6.7 Comparingwith an existingSPR flowdistribution system: (a) integrated flow circuitry (Biacore 1000TM), (b) design of chip flow channels (Biacore A100TM), and (c) combinatory use of channels. Reproduced with permission from [4].",
+        "texts": [
+            " In principle, soft lithography and other small-scale fabrication techniques can be used for all 108 SURFACE PLASMON RESONANCE BIOSENSOR DEVICES alternatives equally. If so, mass production of units should be easy once an assembling fabrication line is setup and automated. The summation of the assessments ranks alternative F as the best and alternativeE as the second best. Thus, this alternative should be pairedwith the sensor chip design according to alternatives C and D. Since these two are probably very similar from the fluidics point of view, any of these can be used with fluidics alternative F. In Figure 6.7, we see examples of this setup principle from the Biacore 1000TM andA100TM systems. It seems that these design solutions coincidewith the conclusion of the mechatronic design analysis, while a more integrated design as in SPREETA is not. However, it must be emphasized that screening should be followed by scoring assessment that would be based on more experimental evidence. This chapter has shown in detail how the mechatronic conceptual design theory is applied. It has shown the benefits of initially defining needs and target specifications, continuing with generating design alternatives based on functions and ranking these"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002832_amm.105-107.448-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002832_amm.105-107.448-Figure1-1.png",
+        "caption": "Fig. 1 Ball bearing model Fig. 2 Rub-impact model",
+        "texts": [
+            " The relation between contact load and deformation of ball is given by: ( ) ( ) 21 1 33 3 0 i e pi i pe e ik k Q \u03b4 \u03b4 \u03b4 \u03c1 \u03c1 = +   = +    \u2211 \u2211 (2) In the equation, \u03b4 is total contact deformation of each rolling element, i\u03c1\u2211 is the sum of main curvature of rolling element and inner ring on the contacting, e\u03c1\u2211 is the sum of main curvature of rolling element and outer ring on the contacting, 0iQ is contact load between inner ring and rolling element at bottom, pik is coefficient of contact deformation of rolling element and inner ring, pek is coefficient of contact deformation of rolling element and outer ring. When under the deformation compatibility conditions, the sum of projection of radial loads of each rolling element at coordinate axis X-Y of radial loads of each rolling element should be equal to external radial load of ball-bearing shown by Fig.1. Following is the expression. 0 0 1 0 0 1 1 1 0 0 2 cos 180 2 cos 180 cos n r i ji j n r i ji ni j n ji ji n i i Q Q Q j n Q Q Q j Q n Q Q \u03d5 \u03d5 \u03d5 \u03d5 \u03b4 \u03d5 \u03b4 = =  = + <    = + + =      = =    \u2211 \u2211 (3) In the expression, j is the number of balls, \u03d5 is the angle between balls. Insert the geometries dimensions of ball bearing 6306,simultaneous equation(2),(3)and solve them, get projection of support reaction of shafting at coordinate axis X-Y signed as xF , yF given by x r y r x F Q y F Q \u03b4 \u03b4  = \u2212   = \u2212  (4) In the equations, x , y are center displacement of shaft journal The Rub-Impact Force Model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001022_j.jsv.2008.05.023-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001022_j.jsv.2008.05.023-Figure6-1.png",
+        "caption": "Fig. 6. A beam element.",
+        "texts": [
+            " 5a where, fBx denotes the contact force and ex denotes the relative displacement at the axis of the bearing. In the numerical analysis, the spring function kx in the axial direction is described by the approximating the linear spring coefficients k1 and k2 (Fig. 5b): kx\u00f0ex\u00de \u00bc k1 0pjexjpa k2 jexj4a ( (14) Based on the experiment, the spring stiffness function is approximated as linear in the lateral direction: fc \u00bc CB _e (15) where _e is the velocity of e and CB is the coefficient of damping. Shafts are modeled using beam element (Fig. 6). Since the torsion stiffness of the shafts in the x direction is high, a small number of motion nodes are enough to express the torsional motion. Thus, a large number of nodes on shafts rotating in x direction are not necessary as in the axial and lateral directions. Here, the motion equation of the beam element e is divided into rotational motion (rotation angles yi x, y j x in x direction) and equation of motions in other directions (displacements u, v, w in x, y, and z directions, and rotation angles yy and yz in the y and z direction, respectively) (Fig. 6). It can be written as follows: Mr e \u20acU r e \u00fe Cr e _U r e \u00fe Kr eU r e \u00bc Fr e (16) ARTICLE IN PRESS Q. Gao et al. / Journal of Sound and Vibration 319 (2009) 463\u2013475 469 where M, C, and K are the mass, damping, and stiffness matrices, respectively. Superscript r denotes the torsional motion of the beam. Ur e is the nodal vector of the rotation angles: Ur e \u00bc \u00bd y i x yj x (17) where subscripts i and j denote nodes i and j: Mb e \u20acU b e \u00fe Cb e _U b e \u00fe Kb eU b e \u00bc Fb e (18) ARTICLE IN PRESS Q. Gao et al"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001872_1.4001203-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001872_1.4001203-Figure3-1.png",
+        "caption": "Fig. 3 R-C-C-R linkage in nonsingular \u201elinks angled\u2026 and singular \u201elinks vertical\u2026 postures",
+        "texts": [
+            " 2, however, the platform velocity is displaced along that line, turning it to remain perpendicular to the connecting link. The immobile version of the four-bar linkage also has a pair of Lie products outside the union span along one dimension, but the Lie products are of opposite signs because the orientation of the follower link has been changed with respect to the mobile version; the resulting roots of the quadratic equation are complex: this linkage is mobile to the first order, immobile to the second order, and hence shaky. 6.2 R-C-C-R Single-Loop Spatial Linkage. Figure 3 depicts an R-C-C-R loop mechanism of a type analyzed by Rico and Ravani 4,28 and by Lerbet and Fayet 23 . This is a spatial four-bar linkage comprised of a pair of R-C dyads. The axes of the revolute R and the cylindrical joint C on the same dyad are parallel to each other, but the common axis direction vector for one dyad is canted relative to the axis direction for the other dyad. A nonsingular posture, which is corresponding to the two link arms being tipped outward in Fig. 3, is summarized in Table 4. The mechanism twist set may be divided into groups of three rows for the three twists of each R-C chain one twist for the revolute joint and two twists for the cylindrical joint . The pair of revolutes in each chain contributes a Lie product, having a velocity component perpendicular to the axes of those revolutes, which are the same as in the preceding examples of the four-bar linkage and the single-link articulation of a moving platform. The prismatic joint adds one translational degree-offreedom normal to the plane, in which that link rotates"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000872_bf00191104-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000872_bf00191104-Figure2-1.png",
+        "caption": "Figure 2.",
+        "texts": [
+            " The velocity of Rayleigh surface waves is less than that of propagation of transversal waves, c2, therefore there arises a question as to how the wedging occurs at velocities higher than the Rayleigh speed. In particular, one may inquire as to the mechanism of wedging in the intermediate range of wedge motion speeds between that of the Rayleigh waves and the velocity of the transversal waves. This problem will now be considered. It is shown that at a super-Rayleigh velocity of wedge motion the scheme of motion becomes essentially different (Fig. 2); no free crack is formed ahead of the wedge, and contact of the wedged body occurs at the wedge head surface only. At some distance from the leading edge of the wedge, the wedged body comes off from the wedge surface and the elastic body surface Int. Journ. ofFractureMech., 8 (1972) 427-434 behind separation points becomes free from stress. Such \"free-streamline flow\" of the wedged elastic body around the wedge is accompanied, in contrast with the sub-Rayleigh case, with the radiation of energy to infinity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000015_bfb0003264-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000015_bfb0003264-Figure6-1.png",
+        "caption": "Fig. 6. Classical bending-torsion flutter instability",
+        "texts": [],
+        "surrounding_texts": [
+            "The pitch-plane model developed above in (13) neglects not only the low frequency phugoid motion, but also higher frequency structural vibration frequencies; thus the aircraft is in fact a \"distributed parameter system\". Even without feedback loops these elastic vibration or \"modes\" can couple together to give rise to \"flutter\" instabilities."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000314_icar.2005.1507392-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000314_icar.2005.1507392-Figure3-1.png",
+        "caption": "Fig. 3. Simulation setup",
+        "texts": [
+            " Number of links is n = 50, which is determined from the requirement for multiple winding around a comparatively small target object with the end-tip part of manipulator in simulations. The other parameters are roughly determined according to preliminary experiments. Figure 2 shows one of those preliminary experimental results. In these experiments, the actuator was controlled roughly and heuristically and the motion data were not measured. The verifiable experiments with precise measurement and control will be developed in the next researches. For the following simulation analyses, initial configurations and locations of the manipulator and target object are as shown in Fig. 3. This manipulator has only one actuated joint at its base and the other joints are all passive ones with the above friction. Let the part from base joint to first joint be called as base link, and the residual part from first joint to end tip be called as HFM part. The length of HFM part is 1.00[m] from the above setting and the length of base link is 0.10[m]. Since the length of base link acts as an offset, actuation of the base link can provide a larger moment to HFM part and control its motion",
+            "0 and resulted in failure. Winding around the target could not be realized by the above sinusoidal actuation. In order to reveal the reason of the above, let us consider the changes of tension in HFM part through the casting motion. Fig. 6 shows the history of load changes at base corresponding to the free-fall motion shown in Fig. 4. Although this load includes the weight of base link itself, the changes in Fig. 6 directly mean those of tension since the base link is represented as a round element as in Fig. 3. Note that in the figure the tension has a maximal peak around t = 0.5[s] where HFM was near to hang-down posture and the variation was not smooth but peaky. This is interpreted as that each link of HFM part passes its lowest point where turning from accelerating to decelerating in order from the base link, and consequently, the end tip takes over the base link when it finally passes its lowest point. The states around such peak point can be considered as those where every links are moving unitedly and comparatively easy to transmit the base motion to end-tip"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000928_2008-01-1019-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000928_2008-01-1019-Figure8-1.png",
+        "caption": "Figure 8: Torque balance at primary flywheel",
+        "texts": [
+            " Both the driveline and the DMF represent energy storage systems, due to their mechanical stiffnesses cj . \u0394E = 1 2 cj(\u03d5i \u2212 \u03d5i\u22121) 2 (18) Here, \u0394\u03d5 = \u03d5i \u2212 \u03d5i\u22121 represents the wind-up or twist angle of various driveline elements including the DMF. \u0394E denotes the energy periodically stored in these springelements. During the periodic energy release processes, this energy acts on the primary flywheel, where the crankshaft position i.e. speed sensor is mounted, introducing additional speed fluctuations. The sensor shown in Figure 8 measures the rotational velocity of the primary flywheel (or alternatively the accessory drive pulley), which is assumed to be rigidly connected to the crankshaft of the engine. The acceleration of the primary flywheel is clearly affected by both the engine torque and the reaction torque due to the driveline or DMF. \u03c9pri(t) = 1 Jpri \u222b t t0 (Teng \u2212 Treact)dt+ \u03c9pri(t0) (19) By measuring the crankshaft alone, it is not possible to accurately estimate the instantaneous engine torque Teng during a dynamic load reaction Treact"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure2-1.png",
+        "caption": "Figure 2. Planetary gear set",
+        "texts": [
+            " The gear contact analysis includes a consideration of the calculated gear mesh misalignment, a non-linear tooth stiffness, as well as the detailed micro-geometry surface definition. The predicted position of the contact on the tooth flank will influence the effect of the gear mesh force. For example, if the contact is away from the centre of the tooth, this will generate a turning moment on the gear, which will have to be reacted by the supporting bearings. The rear planetary gear set in this example is shown in Figure 2. Within the model, the axial load is properly taken by the addition of rolling element thrust bearings (Figure 3), as shown in Figure 4. The clutch and brake parts within the system are modeled as mass, inertia and stiffness components. The completed model is shown in Figure 6, including all gears, bearings, planet carriers and clutches. The boundary conditions of the analysis are then defined. This entails specifying the input speed and torque (or power), and which clutches or brakes are locked"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003859_gt2013-95074-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003859_gt2013-95074-Figure2-1.png",
+        "caption": "Figure 2: Investigated tilting-pad bearing",
+        "texts": [
+            "org/about-asme/terms-of-use 3 Copyright \u00a9 2013 by ASME 3 TEST TILTING-PAD BEARING The investigations were carried out on a five-tilting-pad bearing for power generation applications. The key parameters of the analysis are listed in Table 1. The bearing is mounted in load between pivot position. The oil is supplied to spray-bars between the pads via an outer annular channel. Each spray-bar contains 19 nozzle bores with a diameter of 2.4 mm. The flow rates are further reduced by sealing baffles with a cylindrical bore of a diameter of 502 mm. A drawing of the bearing is shown in Figure 2. The identification of the pad temperatures is enabled by means of 100 thermocouples located 5 mm behind the sliding surface. With reference to Figure 3, the pads feature a radius of R2 = 60 m in axial direction on the back, which enables tilting in axial direction additionally in order to prevent larger influences of misalignment between the bearing and the shaft. Due to the small elliptical area of the contact between the pad and the liner, this pivot is very flexible. The inner radius of the liner is R3 = 322"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002024_j.dam.2010.07.009-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002024_j.dam.2010.07.009-Figure1-1.png",
+        "caption": "Fig. 1. The 3\u00d7 4\u00d7 8 3D chessboard.",
+        "texts": [
+            " We begin by presenting several sufficient conditions for a 3D chessboard not to admit a closed or open generalized knight\u2019s tour (GKT) with given move patterns. Then, we turn our attention to the 3D GKTP with (1, 2, 2) move. First, we show that a chessboard of size L\u00d7M \u00d7 N does not have a closed GKT if either (a) L \u2264 2 or L = 4, or (b) L = 3 and M \u2264 7. Then, we constructively prove that a chessboard of size 3\u00d7 4s\u00d7 4t with s \u2265 2 and t \u2265 2must contain a closed GKT. An L\u00d7M\u00d7N chessboard is a 3-dimension array of cube cells arranged in L rows,M columns andN levels, which is plotted in a x\u2013y\u2013z coordinate system. Fig. 1 presents a 3\u00d7 4\u00d7 8 3D chessboard. Without loss of generality, we assume L \u2264 M \u2264 N . We consider the following move type on 3D chessboards. Suppose the cells of the L \u2264 M \u2264 N chessboard are (i, j, k) where 0 \u2264 i \u2264 L \u2212 1, 0 \u2264 j \u2264 M \u2212 1 and 0 \u2264 k \u2264 N \u2212 1. A move from cell (i, j, k) to cell (r, s, t) is termed an (a, b, c)knight\u2019s move if {|r \u2212 i|, |s\u2212 j|, |t \u2212 k|} = {a, b, c}. For a given (a, b, c)-knight\u2019s move on an L\u00d7M \u00d7 N chessboard, there is associated with it a graph whose vertex set and edge set are {(i, j, k)|0 \u2264 i \u2264 L\u2212 1, 0 \u2264 j \u2264 M \u2212 1, 0 \u2264 k \u2264 N \u2212 1} and {(i, j, k)(r, s, t)|0 \u2264 i, r \u2264 L \u2212 1, 0 \u2264 j, s \u2264 M \u2212 1, 0 \u2264 k, t \u2264 N \u2212 1, {|r \u2212 i|, |s \u2212 j|, |t \u2212 k|} = {a, b, c}} respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure4.32-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure4.32-1.png",
+        "caption": "Figure 4.32 Three pole system root locus",
+        "texts": [
+            " Adding a third pole to the open loop system results in additional asymptotes for the loci because there are now three poles that need a zero as a target. In the first-order system with one pole the single zero target was located at minus infinity along the negative real axis. The two pole system required two zero targets and these are located at 90 degrees to the real axis at \u00b1 infinity. The three pole system requires three zero targets and these are located at minus infinity along the real axis and at angles of \u00b160 degrees to the real axis as shown in Figure 4.32 and in accordance with Rule 3 of root locus construction. The trend is clear. The more lags in the open loop transfer function, the more poles reside in the left half of the \u2018s\u2019 plane and the more aggressively the root loci move into the unstable right half plane. The above examples send an important message. The simple systems with a small number of open loop poles appear to be easy to stabilize. We must Root Locus 151 remember that we often neglect the roots that are outside the frequency range of interest and, while their residues may be small, they still have a small but negative effect on the root loci by forcing themmore quickly to the right and towards the unstable region"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003382_978-1-84996-062-5_5-Figure4.42-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003382_978-1-84996-062-5_5-Figure4.42-1.png",
+        "caption": "Figure 4.42 The cutting layout for tailored blanking [68]",
+        "texts": [
+            " This process is spawning not just a new process but whole new industries dedicated to manufacturing tailored blanks on a just-in-time basis. It is possibly illustrative to look at this one in more detail [68]. A typical car body consists of over 300 pieces, with different thicknesses, treatments and compositions. The cost of assembling them is divided as shown in Figure 4.40. The tailored blank process reduces the cost of material by increasing the yield from the supplied coils from approximately 40 to 65% (see Figure 4.42). It also gives: \u2022 weight reduction; \u2022 lower assembly costs; \u2022 reduced number of parts; \u2022 improved fatigue resistance; \u2022 reduced number of overlapped joints; \u2022 improved corrosion behaviour; \u2022 lower tooling costs; \u2022 improved crash resistance; \u2022 lower press shop costs; and \u2022 fewer sealing operations. The overall saving is measured in dollars per car door alone. Such savings mean that this is the way in which complex sheet metal products are likely to be treated in the future. The process itself is illustrated in Figure 4",
+            " A detailed study made by Toyota Motor Company, who pioneered this process in the 1980s [68], showed the need for low hardness and therefore low-carbon steels and that the weld bead must be at least 80% of the thickness of the butt joint for reasons of strength and deformability. This meant that the gap tolerance was only 0.15mm for 0.8-mm-thick steel for autogenous welds. This limit could be extended by using a filler wire or powder \u2013 this extra material could help with obtaining the correct metallurgy of the weldment, to control the hardness. The material for the pieces of the blank is trimmed from rolls as shown in Figure 4.42. These parts are assembled on a special jig in which, by flexing the sheets, they can be finally pressed together to create a firm compression at the joint line. The assembly is then laser-welded swiftly with a highquality beam, to avoid too much HAZ. During welding the weld bead is monitored by observing the shape on the weld surface using a projected line beam from a He\u2013Ne laser.Theweldment is then ground flat and stacked.The blank stack is then transported to the car assembly plant, where it is pressed to the required shape"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000784_acc.2008.4587070-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000784_acc.2008.4587070-Figure2-1.png",
+        "caption": "Fig. 2. Formation Parameters of Two Neighboring Helicopters",
+        "texts": [
+            " Here, we briefly formulate the parameters of each of these schemes in terms of the helicopters\u2019 state. In order to constraint two helicopters with respect to each other, we try to formulate the relative spatial position of two points named as control point. For each member of the group, we define a control point, which is located on the helicopter\u2019s main axis with an offset of d with respect to the main rotor center. The following vector relation is held for the follower\u2019s control point position (see Fig. 2): pc1 + d1 + l12 + z12 = pc2 + d2 (7) In the above relation, the vectors pc1 and pc2 indicate the position vectors of the leader and the follower helicopters, respectively. The vector l12 + z12 is the relative distance between the leader and the follower control points. The vector di is the relative position of the control point with respect to center of gravity of helicopter i. Representing the sum of the last two terms on the left hand side of the Eq. (7) with p12 and expressing all the terms in the inertial frame, one can write: p12 = R\u22121 If (pc2 \u2212 pc1 + RI2d \u2212 RI1d) (8) In Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003028_glocom.2010.5683528-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003028_glocom.2010.5683528-Figure3-1.png",
+        "caption": "Fig. 3: BCG-TC Phase-II example. After Phase-I, v failed to connect to the original neighbor w in g1 direction. By performing Phase-II, it connects to z.",
+        "texts": [
+            " In fact, when the transmission range is smaller than a certain threshold, the generated topology from Phase-I can be disconnected either due to isolated nodes or multiple components or a combination of both. 2) Phase-II: The second phase of BCG-TC enhances the network connectivity by systematically adding more edges to the nodes having less than k = 4 neighbors. From the connection rule of BCG, each node in a 4-regular BCG has four neighbors in g1, g2, g\u22121 1 and g\u22121 2 directions. In the Phase-II of BCG-TC, if a node is missing a neighbor in any of those directions, the node uses the CTR algorithm to find the next available neighbor in the same direction. For example, in Figure 3, a node v\u2019s neighbor w in g1 direction is out of range. Thus v tries the next neighbor until it finds one in its radio range by computing the w\u2019s neighbor in g1 direction (x = v\u2217g1\u2217g1 = v\u2217g2 1), x\u2019s neighbor in g1 direction (y = v \u2217 g3 1) and so on. When v notices the node z = v \u2217 g4 1 in its radio range, it establishes symmetric link with z only when z completed the Phase-I but has less than k neighbors. Phase-II is summarized in Figure 2. Readers may refer to [14] for details of CTR algorithm"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001155_09544062jmes1190-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001155_09544062jmes1190-Figure2-1.png",
+        "caption": "Fig. 2 Coordinate of a single pad for calculation",
+        "texts": [
+            " Consider the following Reynolds boundary-value problem arising in fluid lubrication of the hydrodynamic bearing with finite length \u03be(p\u0304) = \u03b7 (3) where \u03be(p\u0304) = \u2212 \u2202 \u2202\u03c6 ( h\u03043 \u00b5\u0304 \u2202p\u0304 \u2202\u03c6 ) \u2212 ( d B )2 \u2202 \u2202\u03bb ( h\u03043 \u00b5\u0304 \u2202p\u0304 \u2202\u03bb ) and \u03b7 = \u22123[(x\u0304 cos \u03c6 \u2212 y\u0304 sin \u03c6) + 2(y\u0304 \u2032 cos \u03c6 + x\u0304\u2032 sin \u03c6)] (4) where p\u0304 is the oil film pressure (dimensionless), h\u0304 is the oil film thickness (dimensionless), \u00b5\u0304 is the dynamic viscosity of the oil (dimensionless), d/B is the diameter-to-width ratio of the bearing, \u03c6 is the angle (radian, dimensionless) between the negative direction of the y\u0304 axis and the oil film location, \u03bb is the axial coordinate of calculation (dimensionless), x\u0304 and y\u0304 are the displacements of the centre of the rotor in the x\u0304 and y\u0304 directions (dimensionless), respectively, x\u0304\u2032 and y\u0304 \u2032 are the velocities of the centre of the rotor in the x\u0304 and y\u0304 directions (dimensionless), respectively, \u03b3 is the eccentric angle (radian, dimensionless), \u03d5 is the angle between the connection line of the eccentric angle with the centre of bearing and oil film location (radian, dimensionless), f\u0304x\u0304 and f\u0304y\u0304 are the oil film forces in the negative x\u0304 and y\u0304 directions (dimensionless), respectively, O is the geometric centre of the bearing, Oj is the geometric centre of the rotor, G is the half weight of the rotor (dimensionless), and \u03c9\u0304 is the rotating speed of the rotor (radian, dimensionless), as shown in Fig. 2. There are non-linear problems in engineering applications, which are characterized by variational inequalities involving an elliptic operator and convex sets of admissible functions of obstacle types such as film obstacle problems, elastic contact problems, elastic\u2013plastic torsion problems, fluid lubrication problems, and so on. For fluid lubrication of the finite length journal bearing with the Reynolds boundary, the control equation is valid only in the subdomain / + of the definite domain, whereas the solution must satisfy certain restricted requirements in the subdomain +"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002435_j.triboint.2012.01.011-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002435_j.triboint.2012.01.011-Figure2-1.png",
+        "caption": "Fig. 2. Shaft conditions and corresponding test conditions.",
+        "texts": [
+            " Airflow in the gap space did not affect grease movement, which was checked using powdered milk instead of grease. Thus, any grease observed seeping into the gap would mean that the mechanisms of grease leakage did not entail centrifugal force or the pushing-out force generated by the rolling elements or cage motion. The disk and glass plate were set in the three positions shown in Fig. 1(a) and (b): the normal position (CASE 1), an inclined disk that leads to disk precession motion (CASE 2), and an inclined glass plate (CASE 3). As shown in Fig. 2, these positions correspond to the correct bearing and shaft positions, shaft inclination against the bearing, and shaft precession motion in the actual bearing-and-shaft assembly, respectively. Tables 1 and 2 list the test greases and test conditions, respectively. No. 3 PFPE grease was applied for CASE 1 and No. 4 for CASES 2 and 3, in considering the assumed actual application of each type of grease. The disk rotation speed for CASE 2 was lower than that for CASE 1, in order to avoid vibration due to imbalanced mass"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002739_1.4733331-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002739_1.4733331-Figure1-1.png",
+        "caption": "FIG. 1. A slanted fiber contacts with a rigid smooth surface. Both the undeformed configuration and deformed configuration of the fiber are shown.",
+        "texts": [
+            " We found that the fiber\u2019s detachment mode is affected by its bending stiffness, slanted angle and the applied shear force. The normal pull-off force increases with the applied shear force. By using the stability criterion of energy minimization, we also prove that all the adhesion equilibrium states are stable. The theoretical results are expected to be useful for interpreting experimental results of frictional adhesion test with similar microstructures, and to be valuable for optimum designs of fibrillar dry adhesives. As shown in Fig. 1, we consider a single straight slanted fiber which is in side contact with a rigid smooth surface, for a large enough preload is applied in normal direction. The fiber\u2019s length is l, the slanted angle is a and bending stiffness is EI with Young\u2019s modulus E and inertia moment I. Here, we assume the fiber\u2019s aspect ratio is high and the axial strain is comparatively small, so the inextensible beam theory can be adopted to describe the deformation of the fiber, which is the same as that done by Majidi et al",
+            " Possible adhesion between the top surface of the fiber and the surface in the detachment process is also neglected, in that our main interest is exploring the directional adhesion behavior of the fiber which is in side contact with the contacting surface and loaded in force-controlled mode rather than displacement-controlled mode. Both a normal force Fn and a shear force Ft as well as the surface force cause the fiber to deform, thus varying the contacting length l a where a is termed the \u201cnon-contact length\u201d of the fiber. The total free energy of the system U contains three parts: the bending energy of the fiber UB, the surface energy US and the potential of external force UP. Using the convected coordinate \u00f0s; h\u00de as sketched in Fig. 1, the expressions of these energies are as follows:19,20 UB \u00bc \u00f0a 0 1 2 EI @h @s 2 ds; US \u00bc \u00f0l a\u00dex; UP \u00bc Ft\u00bd\u00f0l a\u00de \u00fe \u00f0a 0 sin hds \u00fe Fn \u00f0a 0 cos hds; (1) where x denotes the equivalent work of adhesion per unit length for the fiber. At the state of equilibrium, the total free energy of the system reaches its minimum, requiring that the variational derivatives of U with respect to h and a vanish. Applying a variation method to Eq. (1), we can obtain the equilibrium equation of the system, EI @2h @s2 \u00fe Fn sin h\u00fe Ft cos h \u00bc 0; 0 < s < a; 1 2 EI @h @s 2 \u00bc x; s \u00bc a; (2) with natural boundary conditions h\u00f00\u00de \u00bc a; h\u00f0a\u00de \u00bc p 2 : (3) The derived Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002511_pedstc.2012.6183324-FigureI-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002511_pedstc.2012.6183324-FigureI-1.png",
+        "caption": "Figure I. Vehicle system diagram propolled by a LIM",
+        "texts": [
+            " A coefficient that is reversely dependent to speed is determined to describe the effects that the LIM speed causes in the magnetization branch of the equivalent electrical circuit [1], [4]. In this paper we first express the dynamic model of the linear induction motor with end effects included. Then, we describe the indirect vector control of LIM. Finally, in this work a simulation model of LIM with and without the end effect is developed following the equivalent circuit model. Simulation results show the validity of the indirect vector control of LIM. II. DYNAMIC MODEL OF LINEAR INDUCTION MOTOR A conceptual construction of a LIM that is used in linear metro is depicted in Fig.I. The short primary is movable and the infinite secondary is fixed and flatted on the railway track. In a LIM, the primary is similar to the stator of a rotary induction motor (RIM), is hanged below the redirector, which is supplied by the inverter on the vehicle. The secondary consists of a sheet conductor with a back iron to return path for the magnetic flux. Thus, the primary voltage or current excitation produces a magnetic field from the front to the back IEEE Catalog Number: CFP121IJ-ART 193 ISBN: 978-1-4673-0113-8/12/$31"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001634_978-3-642-00644-9_19-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001634_978-3-642-00644-9_19-Figure1-1.png",
+        "caption": "Fig. 1 (a) Photo of a Kobot. (b) Top-view of a Kobot sketch showing the body (circle), the IR sensors(small numbered rectangles), and the two wheels (gray rectangles). (c) The reference frame is fixed to the center of the robot where the x-axis coincides with the rotation axis of the wheels. The forward velocity (u) is along the y-axis. \u03c9 denotes the angular velocity of the robot. The y-axis of the bodyfixed reference frame makes an angle of \u03b8 with the sensed North direction (ns) at the instant the figure is drawn, which is the current heading of the robot. With kind permission from Springer Science+Business Media: Swarm Intelligence, Selforganized flocking in mobile robot swarms, volume 2, number 2-3, 2008, Ali Emre Turgut, Hande C\u0327elikkanat, Fatih Go\u0308kc\u0327e and Erol S\u0327ahin, Fig. 1",
+        "texts": [
+            " The informed robots do not signal that they are \u201cinformed\u201d, and instead guide the rest of the swarm by their tendency to move in the preferred direction. We present experimental results on both physical and simulated robots, and show that the self-organized flocking of a swarm of robots can be effectively guided by a minority of informed robots within the flock, without an explicit leadership mechanism. Then we analyze the system\u2019s performance under various conditions. In this study, we use Kobot and its physics-based simulator CoSS [15]. Kobot is a light-weight (12 cm diameter), differentially driven robotic platform (Fig. 1(a)). It has two main sensory systems: the Infrared Short-Range Sensing System (IRSS) and the Virtual Heading Sensor (VHS). IRSS is composed of 8 infrared sensor modules located at 45\u25e6 intervals around the base (Fig. 1(b)), and is used for short-range proximity measurements. It uses modulated infrared signals to minimize the environmental interference, and crosstalk among robots. The sensors can detect objects within a 21 cm range at seven discrete proximity levels, and can distinguish kin-robots from obstacles. The VHS is used for virtually \u201csensing\u201d the relative headings of the neighboring robots. It consists of a digital compass and a wireless communication module. The robot\u2019s heading with respect to the sensed North is measured using the compass module and is broadcasted to other robots through wireless communication",
+            " The heading alignment vector h tries to align the robot with its neighbors, and is calculated as: h = \u2211 j\u2208N ei\u03b8j \u2016\u2211 j\u2208N ei\u03b8j\u2016 where N denotes the set of VHS neighbors, \u03b8j is the heading of the jth neighbor converted to the body-fixed reference frame and \u2016 \u00b7 \u2016 calculates the Euclidean norm. The proximal control behavior is responsible for flock cohesion and collision avoidance. The normalized proximal control vector p is calculated using the infrared readings from the IRSS. It is a vector sum of virtual forces which are assumed to act on each infrared sensor: p = 1 8 \u2211 k fkei\u03c6k where k \u2208 {0, 1, \u00b7 \u00b7 \u00b7 , 7} denotes the sensor positioned at angle \u03c6k = \u03c0 4 k with respect to the x-axis (see Fig. 1(b)), and fk denote the virtual force acting on the sensor. The virtual forces are taken to be proportional to the square of the difference between the current detection level (ok) of the sensor, and the desired detection level (odes). The desired detection level is defined as an intermediate detection level (3) if the sensed object is another robot and 0 if it is an obstacle. This setting motivates the robot to keep at an optimal distance from its peers and escape from obstacles. fk is then calculated as follows: fk = { \u2212 (ok\u2212odes) 2 C if ok \u2265 odes (ok\u2212odes) 2 C otherwise. The direction preference vector d is calculated as: d = dp \u2212 ac where ac is the current heading vector of the robot coincident with the yaxis of the body-fixed reference frame (see Fig. 1(c)), and dp stands for the preferred direction. The desired heading vector, a, is used to calculate the forward (u) and angular (\u03c9) velocities. u is calculated via diminishing the maximum speed umax according to the robot\u2019s momentary urge to turn. This urgency is given by the difference between the desired heading and the current heading of the robot, calculated by a dot product of the two vectors: u = { (a \u00b7 ac) umax if a \u00b7 ac \u2265 0 0 otherwise The angular velocity \u03c9 is controlled by a proportional controller: \u03c9 = 1 2 ( ac \u2212 a) We evaluate the performance of the system using two different metrics"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000662_j.jmatprotec.2007.11.038-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000662_j.jmatprotec.2007.11.038-Figure3-1.png",
+        "caption": "Fig. 3 \u2013 Simulation of straight tooth cylindrical gear generation w movement a, feeds b, c and turns \u03d5, \u03d51 and \u03d52.",
+        "texts": [
+            " 2)} move tool {tool shift corresponding to a wheel revolution} n g t 334 j o u r n a l o f m a t e r i a l s p r o c e s s i 0, 0 base point of the shift b, 0 {value of the shift (Fig. 2)} copy tool {tool copy and its placement on the turning diameter of the wheel dt} 0, 0 {base point} 0, a {location of the tool copy (Fig. 2)} subtract gear wheel tool {subtraction of the dipped tool solid volume from the gear wheel model} rscript {restarting the procedure}. The kinematics of the hobbing and chiselling simulation was presented in Fig. 3. In the case of hobbing (Fig. 3a), the tool is a model of a plain milling cutter. The computer simulation of the gear generation takes place in a three-dimensional environment. A tooth profile is formed due to the turning of the gear wheel and the tool. The helix results from the hob blade moving along the wheel axis. The feed motion of the tool b is dependent on the rotary motion of the generated gear wheel \u03d5 in accordance with the transmission ratio of the technological gear, Eq. (3). Shift c corresponds to the value of the axial feed per wheel turn. A computer simulation of gear generation by chiselling was carried out in a similar way (Fig. 3b). The tool in this case is a solid model of a pinion cutter. The turn of the e c h n o l o g y 2 0 4 ( 2 0 0 8 ) 331\u2013342 tool by angle \u03d51 is related to the wheel turn \u03d52, in accordance with the transmission ratio of the technological gear. Examples of the flanks of the teeth obtained through hobbing and chiselling simulation is presented in Fig. 4. In the first case the flanks of the toothed wheel are composed of a number or regularly distributed surfaces corresponding to the successive layers of the material being removed by the tool blades"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001727_0022-2569(69)90006-8-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001727_0022-2569(69)90006-8-Figure6-1.png",
+        "caption": "Figure 6. R -C-C-R mechanism.",
+        "texts": [
+            "5) reduce to: (at z --+ az 3) = (a34 -4-a41 ) $2 sin(cc23 --- 7z34) = ---Sx t sin ~4t $3 = 0 -T-S+ sin(z~z3 -+-~3+) = ___St1 sin ~tz The links a23 and a34 are thus coUinear and parallel to the links a~2 and a+~. If S ~ t = 0 then all links become collinear. For double change points the sine and cosine laws for spatial triangles give the geometry of the Bennett mechanism. 4. Classification of the R - C - C - R Mechanism R - C - C - R mechanisms have been reported by Dimentberg and Al tman [6]. The eight dual angles for the most general form (Fig. 6) are, C212 = 0 +~a12 0 1 = 0 t +8S11 ~23 = 0~--[- 8a23 0 2 - 0 2 + e S 2 c23~,, = 0 + ca34 03 = 03--beS 3 c2+1 = ~ + ea4t 04=0++eS4+ (4.1) (i) Displacement equations By substituting (4.1) into equations (2.1)-(2.5) and separat ing into p r imary and dual components we obtain: a34 cos 0+ = (a23 -- a+t ) - a I 2 cos 01 sin 02 =s in 0t $2 sin ~ cos 0z = a3+(cos 01 sin 0+ + c o s ~ sin 0t cos 04) - ( a z 3 - a+l)cos ~ sin 01 + S i t sin = cos 0t sin 03 = sin 04 269 $3 sin 7 cos 83 =a l z ( s in 81 cos 8++ cos m cos 8 t sin 84) - (a23 - a 4 t ) cOs ~t sin 83 + $44 sin ~ cos 04 al 2 cos 82 = a34 cos 8+-- (a23 -- a41 ) at z cos 8 t = a34 cos 83 + (a 23 - a+ t) (4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000923_bfb0110378-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000923_bfb0110378-Figure5-1.png",
+        "caption": "Fig. 5. Towed cable system and finite link approximate model.",
+        "texts": [
+            " E x a m p l e 14 ( T o w e d cab le s y s t e m ) Consider the dynamics of a system consisting of an aircraft flying in a circular pattern while towing a cable with a tow body (drogue) attached at the bottom. Under suitable conditions, the cable reaches a relative equilibrium in which the cable maintains its shape as it rotates. By choosing the parameters of the system appropriately, it is possible to make the radius at the bottom of the cable much smaller than the radius at the top of the cable. This is illustrated in Figure 5. The motion of the towed cable system can be approximately represented using a finite element model in which segments of the cable are replaced by rigid links connected by spherical joints. The forces acting on the segment (tension, aerodynamic drag and gravity) are lumped and applied at the end of each rigid link. In addition to the forces on the cable, we must also consider the forces on the drogue and the towplane. The drogue is modeled as a sphere and essentially acts as a mass attached to the last link of the cable, so that the forces acting on it are included in the cable dynamics"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure2-1.png",
+        "caption": "Figure 2. Determination of target position and target direction.",
+        "texts": [
+            " *Corresponding author. e-mail: lkb7194@korea.ac.kr Once three points are identified on the world coordinate system, only one circle of radius R with its center at C, exists that passes through these three points. From Figure 1(b), we can see that the vehicle\u2019s current position is not on the virtual midline and furthermore, that its direction is not parallel to the tangential direction at point p on the virtual mid-line. Therefore, the indications are that the vehicle is not following the virtual midline properly. Figure 2 illustrates how to determine target position and target direction. The target position is defined as a point on the virtual mid line which is separated from the current vehicle position by the look-ahead distance in the Yw direction. This look ahead-distance varies depending upon the driver\u2019s driving habits or tastes, and normally ranges from 15 to 25 meters (Raksincharoensak, 2004). A target direction is determined as the tangential line at the target position on the virtual mid line. for a vehicle to follow a virtual mid line in a safe way, its current position should be on the virtual mid line and furthermore, its direction should be tangential to the virtual mid-line"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000812_tia.2006.886430-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000812_tia.2006.886430-Figure2-1.png",
+        "caption": "Fig. 2. Double-motor drive configurations. (a) Seven-leg converter with one boost (DM-1B-7L) and seven-leg converter without boost (DM-0B-7L) (f replaced by a short circuit). (b) Six-leg converter with one boost (DM-1B-6L) and six-leg converter without boost (DM-0B-6L) (f replaced by a short circuit).",
+        "texts": [
+            " 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/TIA.2006.887430 and do not use boost inductor filters have been proposed in [7] and [8]. This paper proposes eight drive configurations that operate with reduced number of switches and boost inductor filters, but addressed to applications employing a three-phase main supply. The ac drive systems proposed are shown in Figs. 1 and 2. Four of the configurations are single-motor drives (Fig. 1), and the other configurations are double-motor drives (Fig. 2). The drives provide both bidirectional power flow and powerfactor control. Besides using a reduced number of boost inductor filters, the configurations use fewer switches than the equivalent conventional configurations. As an example, the five-leg [Fig. 1(a)] and the four-leg [Fig. 1(b)] configurations use fewer switches than the six-leg converter [13] [Fig. 3(a)] and the five-leg converter [Fig. 3(b)], respectively. Four conventional configurations are compared to the proposed configurations. The conventional single-motor drive configurations are presented in Fig. 3, and the double-motor-drive configurations are obtained by connecting to the dc-link a second motor supplied by another inverter. The proposed configurations are indicated by the following abbreviations: 1) SM-2B-5L: single-motor, two-boost, five-leg [Fig. 1(a)]; 2) SM-1B-5L: single-motor, one-boost, five-leg [Fig. 1(a)]; 3) SM-2B-4L: single-motor, two-boost, four-leg [Fig. 1(b)]; 4) SM-1B-4L: single-motor, one-boost, four-leg [Fig. 1(b)]; 5) DM-1B-7L:double-motor,one-boost,seven-leg[Fig. 2(a)]; 6) DM-0B-7L: double-motor, no-boost, seven-leg [Fig. 2(a)]; 7) DM-1B-6L: double-motor, one-boost, six-leg [Fig. 2(b)]; 8) DM-0B-6L: double-motor, no-boost, six-leg [Fig. 2(b)]. The conventional configurations are referred by the following abbreviations: 1) SM-6L: single-motor, six-leg [Fig. 3(a)]; 2) SM-5L: single-motor, five-leg [Fig. 3(b)]; 3) DM-9L: double-motor, nine-leg [Fig. 3(a) with an addi- tional inverter-motor system]; 4) DM-8L: double-motor, eight-leg [Fig. 3(b) with an addi- tional inverter-motor system]. The machines used in this paper are typical three-phase machines. Adopting a fixed coordinate reference frame, the 0093-9994/$25.00 \u00a9 2007 IEEE mathematical model that describes the dynamic behavior of the three-phase induction machine is given by [14] vsdq = rsisdq + d dt \u03bbsdq (1) vrdq = rrirdq + d dt \u03bbrdq \u2212 j\u03c9r\u03bbrdq (2) \u03bbsdq = lsisdq + lsrirdq (3) \u03bbrdq = lsrisdq + lrirdq (4) vso = rsiso + lls d dt iso (5) vro = rriro + llr d dt iro (6) Te =Plsr(isqird \u2212 isdirq) (7) where vsdq = vsd + jvsq, isdq = isd + jisq , and \u03bbsdq = \u03bbsd + j\u03bbsq are the voltage, current, and flux dq vectors of the stator, respectively, vso and iso are the homopolar voltage and current of the stator, respectively (the equivalent rotor variables are obtained by replacing the subscript s by r), Te is the electromagnetic torque, \u03c9r is the angular frequency of the rotor, rs and rr are the stator and rotor resistances, ls, lls, lr, and llr are the self- and leakage inductance of the stator and rotor, respectively, lsr is the mutual inductance, and P is the number of pair of poles of the machine",
+            " 1(b) with the boost f2 replaced by a short circuit. All previous voltage relations (36)\u2013(42) and (32) and (34) are valid except for v50, which is given by v50 = 0. Relations (28)\u2013(31) are also valid for this case. The grid model is depicted in Fig. 4(b). However, in this case, besides the small ac current due to the filter unbalance, an extra ac current, approximately equal to ig2/2, traverses the dc-link capacitor. This current is due to the connection of the phase 2 of the grid to the capacitor midpoint. The models for topologies shown in Fig. 2 can be derived as done for topologies with a single motor. The models for the double-motor systems are presented in the Appendix I. The motor torque control, which includes the flux control, can be accomplished by controlling either the dq currents (as in field-oriented control) or the dq voltages (as in volts/hertz control). The grid power-factor control is generally achieved by controlling the grid currents. Initially, consider that both torque and power-factor control should be accomplished by controlling just the currents"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002737_s00359-012-0714-5-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002737_s00359-012-0714-5-Figure7-1.png",
+        "caption": "Fig. 7 Calculation of expected errors in home vectors in a 3D path integration task, depending on an assumed misjudgement of inclinations. The inset shows the experimental situation as used by Grah et al. 2005. Ants that were transferred from the feeder (F) to a test field would return to the fictive nest position on the stippled path, with an angle b relative to the first leg of the training course. For different inclinations (a) of the \u2018\u2018hill\u2019\u2019 in this course, we assumed that a is increased by a constant error Du (both on ascent and descent), leading to an error in the correct home vector (b). The error of b is shown for two values of Du, as observed in experiments I and II",
+        "texts": [
+            " From the inclined path segments and their slopes, desert ants derive an estimate of the projection of these segments onto the ground and continually feed these estimates as distance information into their two-dimensional path integration system, with remarkable precision (Grah et al. 2005). In order to explore the impact that accuracy of slope discrimination (Du) may have on path integration in undulating terrain, we turn back to the 3D experiment performed by Grah et al. (2005). In this study, ants were trained to walk in a channel system for 3 m in horizontal direction, then the channel made a 90 turn and the ant had to climb a 70 slope for 2 m, from the top they descended a 2 m ramp, again with a 70 slope, to reach the feeder located at the ground (see sketch in Fig. 7). From the feeder, individual ants were taken and the homing direction (b) was recorded on a distant test field (see Figs. 1, 2, 3 in Grah et al. 2005). The principle of this experiment was to translate the ant\u2019s (ground) distance measurement into a directional decision on the test field, where the homing direction can be recorded with high precision. A systematic deviation from the correct ground distance estimate would become evident in this paradigm as a deviation in the homing angle (b). In the experiment reported by Grah et al",
+            "1 ), demonstrating an accurate calculation of the ground distance in the 3D paradigm. We now explored as to what impact an inaccurate slope detection would exert on the path integration performance. By assuming that an ant misinterprets the slope by a constant angle, in the range of Du as determined in Fig. 6c, we can estimate how far her homing direction on the test field would deviate from the correct vector direction b due to this systematic error in slope estimation. The calculated error angles are shown in Fig. 7, for a value of minimal Du = 12.5 , as found in experiment II (for comparison, the values for a Du = 25 are shown as well). Remarkably, a 12.5 error in slope measurement caused only a moderate error in the homing vector, less than 7 deviation of b for training inclinations up to a slope a = 45 . Only for a 60 or 75 training slope, the home vector would deviate by more than 10 . The calculations shown in Fig. 7 are important since they suggest that there might be no strong selective pressure enforcing a further increase in the accuracy of slope discrimination to values smaller than Du = 10 . Thus, although the Du of *12 found in experiment II may not constitute the true minimally distinguishable slope differences (see above), it seems a reasonable approximation. For the 60 training inclination, however, the behavioural Du was larger, and could not be reduced by the avoidance training. Thus, the deviations in the home vector b would be larger than calculated in Fig. 7. Nonetheless, it seems not likely that ants are completely unable to discriminate all slopes larger than say 55 : in the experiment of Grah et al. (see sketch in Fig. 7), the training slope was 70 . If the ants had perceptually lumped together all slopes above 55 , we would expect a deviation of the mean vector from the ground control and a much higher variance. Although the circular standard deviation in the 3D test was higher than in the ground control (21.4 vs. 14.3 ), the difference is not extreme and in particular the mean vectors were virtually identical (25.1 and 24.3 , respectively; Grah et al. 2005). In summary, C. fortis foragers were able to distinguish slopes that differed by little more than 10 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001618_s0580-9517(08)70600-7-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001618_s0580-9517(08)70600-7-Figure2-1.png",
+        "caption": "FIG. 2.",
+        "texts": [
+            " ALLEN bacterial suspension or the inoculated culture medium and the substrate under investigation, and in the cathode compartment a solution of a reducible species, e.g., potassium ferricyanide, so as to establish the cathode as a nonpolarizable counter-electrode. The side arm of the anode half-cell is used for introduction of substrate solutions, anti-metabolites, etc., after the electrochemical cell has been assembled and equilibrated to bath temperature. The only differences between the cells illustrated, is that the anode compartment of the mini-electrochemical cell (Fig. 2) can be used with quantities of liquid from 25 to 30m1, and because of the design, the electrochemically active species has a shorter path to travel from the bulk of the solution to the electrode surface (the electron sink). As a result of the latter, a greater efficiency is obtained from the system. Whereas in potential measurements it is not necessary to completely immerse the working electrode in the liquid being studied, with current measurements this is a necessity in order to (1) obtain the maximum current output for the size of the electrode used, and (2) achieve the maximum degree of reproducibility of results",
+            " These cells continued to metabolize a useable substrate until it was completely depleted or until the waste products of metabolism accumulated and the conditions necessary for development were no longer present. An explanation for the charge-discharge phenomen (spur) observed in the ascending portion of the current/time curve is ascribed to a transient steady state which manifests itself again as the second plateau in the descending section of the curve. Because of the need to use smaller quantities of substrates and microorganisms the cell shown in Fig. 2 was constructed. In designing this electro-chemical cell another objective was also taken into consideration, that is, increased electrochemical efficiency by reducing the distance the bacteria. had to travel to reach the electrode surface from the bulk of the solution. With this cell it was found that the anolyte could consist of 10 ml of the bacterial suspension (section 111-E) diluted with 10 ml of the pH 6.7 phosphate buffer. T o the anolyte is then added 5 ml of 0.0473 M of an active substrate and coulombic measurements made as previously described"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003971_icra.2012.6225012-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003971_icra.2012.6225012-Figure3-1.png",
+        "caption": "Fig. 3: The bended bar used to obtain the results in Table II",
+        "texts": [
+            " For this purpose, we have conducted a set of experiments where we consider the static deflections for bars of different lengths. The mass per unit length of the bar was measured to A\u03c1 = 0.64. We have estimated the value of IcG to be approximately IcG= 100. The measured results are presented in Table II together with the resulting model deflections using IcG = 100. As can be seen, the accuracy of both the model deflection and the RobWork deflection is within 1\u22122cm of the actual deflection of up to more than 30cm. To validate the model also when having bended the bar, we studied the bar shown in Figure 3. Results are shown in Table III. Again, we notice that the deviation between model and reality is within a few centimeters, but somewhat larger than for the unbended bar. We will conduct further experimental studies to both adjust the model and find possible reasons for the discrepancy. D. Vibrational analysis As an independent validation, we now study the vibrational movements of the bar. We use a damped beam model, where one end is fixed and gravitational load and damping is neglected. We may then write Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000393_9780470061565.hbb032-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000393_9780470061565.hbb032-Figure4-1.png",
+        "caption": "Figure 4. Schematic of an immunochromatographic electrochemical biosensor.4 [Reproduced from Lu et al.4 with permission from Royal Society of Chemistry.]",
+        "texts": [
+            " The concentration of an analyte at the Hg electrode, for td (duration of accumulation), nh (number of moles of electrons transferred in the half reaction), il (limiting current during reduction of the metal) and VHg (volume of the electrode), CHg, is given by: CHg = i1td nhFVHg (5) The expression for current produced by anodic stripping depends on the particular type of Hg electrode but is directly proportional to the concentration of the analyte at the electrode. The main advantage of stripping analysis is the preconcentration of the analyte onto the electrode before making the actual current measurement. Anodic stripping can achieve detection of concentrations as low as 10\u221210 M. Lu et al. have reported the use of ASV technique for detection of immunochromatographics such as human chorionic gonadotropin (HCG) at built-in single-use screen-printed electrodes (Figure 4).4 Wang et al. have reported the immobilization of single-stranded DNA on gold colloid particles associated with a cysteamine monolayer on gold electrode surface.5 They have studied the hybridization of a silver nanoparticle\u2013oligonucleotide DNA probe, followed by the release of silver metal atoms anchored on the hybrids by oxidative metal dissolution, and the indirect determination of the released solubilized AgI ions by ASV at a carbon fiber microelectrode. The studies show good correlation for DNA detection in the range of 10\u2013800 pmol l\u22121 and allow a detection level as low as 5 pmol l\u22121 of the target oligonucleotides"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000284_acc.2006.1657378-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000284_acc.2006.1657378-Figure5-1.png",
+        "caption": "Fig. 5. Master r and \u03c6 directions",
+        "texts": [],
+        "surrounding_texts": [
+            "To cope with the time-varying delay, two approaches are possible: a) the estimation of the plant state that was explained in section II. C and, b) the addition of a buffer d [3]. The buffer may be used to save the information that arrives from the opposite side of the teleoperation loop during a time that exceeds the maximum time delay. This information is then feed into the controllers at a constant rate to each controller. By using this method, the time delay may be kept constant at the expense of making it larger. For the teleoperation experiment, we implemented the buffer idea. However, our control law is capable of producing acceptable performance of the NCS even in the absence of the buffer, and without any other time-varying delay compensation scheme, as shown in Figures 7, and 8, which show the tracking performance of the remote slave robot. The force reflected to the master in this experiment is due to the dynamics of the slave robot. In other words, even though there is no force applied to the robot from an obstacle, gravity, friction and time delays force the robot to have a settling time different from zero. This in turn produces an error between the references sent to the slave and the actual state measurements, which forms the basis for the force reflection control laws designed in equations 5-8. The buffer size was chosen of 20, so this imply a constant time-delay in the loop of almost 20 times the average delay. This size was chosen assuming that neither the network nor the computer processing time will induced any longer delay and it worked reasonable for the experiment. Since the delay time in the loop was incremented by the buffer inclusion, the control gains were tuned again. In the case of the robot control law , the gains were as follows: Krv = 100.0, B\u03c6\u03b8 = 1\u00d7105 and K\u03c6\u03b8 = 2.5\u00d7106. For the haptic device control law, the gains were: Br = 0.5, Kr = 0.001, Krv = 100.0, B\u03c6 = 0.1, B\u03c6\u03b8 = 100.0, K\u03c6\u03b8 = 2500.0 and \u03bb = 0.04. The experimental obtained by using the buffer are shown in Figures 9 to 12. From these results we see that the tracking in velocity and angle experience a longer delay, caused by the buffer. However, the tracking in the angle is more accurate than when the buffer was not used."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure7-1.png",
+        "caption": "Fig. 7. Structure of the riveted joint model",
+        "texts": [
+            " On the upper skin there are some disturbances connected with local buckling, but there is no significant influence of this phenomenon on the lower skin. Fragment of the bottom skin, near rib no 21 (fig. 6) was chosen for local I level analysis. Methods for FEM analysis of riveted joints of thin walled aircraft structures 947 The Riveted joint FEM model was built for this region. The Presence of the rivet was taken into account, as well as the distance between middle surfaces of jointed parts. The structure of the model is presented in fig. 7. Dimensions of the model are 400 x 150 mm. Shell elements (Quad4, Tria3) were used. For rivets in the middle, where one can see holes, model presented in fig.8 was used. The model is based on previous work [12,13]. It consists of two circular surfaces, with rivet diameter, connected with rigid MPC element. On the surfaces\u2019 edges, on sheets plans there are GAP contact elements. They model interaction between the rivet shank and sheets. Manufactured head and driven head are not represented. Instead of them displacements in axial direction (z) of nodes of sheets and rivet, on hole edges were tied with MPC elements"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001104_09544062jmes817-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001104_09544062jmes817-Figure1-1.png",
+        "caption": "Fig. 1 Loads and displacement of the angular ball bearing",
+        "texts": [
+            " Then, the dynamic behaviours of the rotor system supported by an angular ball bearing is investigated under the effects of different parameters, and the dynamic characteristics of ball bearings are computed considering the vibration of the rotor system. The study will provide evidence for accurately analysing the characteristics of the ball bearing and reasonably choosing the bearing operation parameters. In order to investigate the dynamic characteristics of a rotor ball bearing system, the non-linear contact forces should be determined. Figure 1 illustrates the relationship between the load {Fx , Fy , Fz , My , Mz} and displacement of an angular ball bearing {X2, Y2, Z2, \u03b8y , \u03b8z}. The geometric model of a ball and races is shown in Fig. 2. O is the mass centre and geometric centre of the outer ring, O2 is the mass centre and geometric centre of the inner ring, Ob is the mass centre and geometric centre of the ball, which are prescribed, respectively, in the inertial frame (X , Y , Z), the inner ring fixed frame (x, y, z), and the ball fixed frame (x, y, z)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000839_00029890.2008.11920595-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000839_00029890.2008.11920595-Figure2-1.png",
+        "caption": "Figure 2. Trajectories Ys(x), Yw2(x), Yw3(x), Yav(x) (labeled A), and Ye(x) (dashed), for terminal speed 40 m/s and initial speed 40 m/s at launch angle 60\u25e6.",
+        "texts": [
+            " So a better trajectory approximation seems to be Yav(x) = [Ys(x) + Yw2(x)]/2, which gives excellent range estimates with errors of less than 1% in all test cases except those having sI \u2208 {40, 80} m/s with \u03c6 \u2208 {60\u25e6, 75\u25e6}. A higher-order approximation, not pursued in [15], is found by retaining only terms of degree 3 or less in (21); this leads to a messy expression, Yw3(x), omitted here but given in [7]. Yw3(x), like Yav(x), gives more accurate range estimates than both Ys(x) and Yw2(x) in all test cases, but Yav(x) generally gives the best estimates. More details on test case results for various approximations are provided in [7]. Figure 2 depicts trajectories Ys(x), Yw2(x), Yw3(x), Yav(x), and Ye(x) for one test case. Rewriting (1) in terms of derivatives of y, s, and \u03b8 with respect to x and repeatedly differentiating the result, as in Littlewood [11], generates Ye(x) = b a x \u2212 g 2a2 x2 \u2212 cgsI 3a3 x3 \u2212 cg 2cs2 I \u2212 bg/sI 12a4 x4 + O(x5), (23) for sufficiently small x . Although (23) has some value as a trajectory approximation (when extended to error O(x7), its range estimates on the test cases are generally a bit worse, and less consistent, than those of Yav(x)), it is most useful in proofs (as in [11]) and for deducing other results, such as this expansion for the range x : x = 2a2 g \u03c6 \u2212 8a4c 3g2 \u03c62 + 40a6c2 + 6a2g2 9g3 \u03c63 + O(\u03c64), (24) for sufficiently small \u03c6, which agrees with Galileo\u2019s range 2a2g\u22121 tan \u03c6 when c = 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.1-1.png",
+        "caption": "Fig. 3.1. Characterization of generic Link i of a manipulator",
+        "texts": [
+            " Further, denoting the vector R0 1p 1 by r0 1, it is p\u03070 = o\u03070 1 +R0 1p\u0307 1 + \u03c90 1 \u00d7 r0 1 (3.14) which is the known form of the velocity composition rule. Notice that, if p1 is fixed in Frame 1, then it is p\u03070 = o\u03070 1 + \u03c90 1 \u00d7 r0 1 (3.15) since p\u03071 = 0. Consider the generic Link i of a manipulator with an open kinematic chain. According to the Denavit\u2013Hartenberg convention adopted in the previous chapter, Link i connects Joints i and i + 1; Frame i is attached to Link i and has origin along Joint i+1 axis, while Frame i\u22121 has origin along Joint i axis (Fig. 3.1). Let pi\u22121 and pi be the position vectors of the origins of Frames i\u22121 and i, respectively. Also, let ri\u22121 i\u22121,i denote the position of the origin of Frame i with respect to Frame i\u2212 1 expressed in Frame i\u2212 1. According to the coordinate transformation (3.12), one can write1 pi = pi\u22121 +Ri\u22121r i\u22121 i\u22121,i. Then, by virtue of (3.14), it is p\u0307i = p\u0307i\u22121 +Ri\u22121r\u0307 i\u22121 i\u22121,i + \u03c9i\u22121 \u00d7Ri\u22121r i\u22121 i\u22121,i = p\u0307i\u22121 + vi\u22121,i + \u03c9i\u22121 \u00d7 ri\u22121,i (3.16) which gives the expression of the linear velocity of Link i as a function of the translational and rotational velocities of Link i\u2212 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003859_gt2013-95074-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003859_gt2013-95074-Figure5-1.png",
+        "caption": "Figure 5: Technical drawing of the unbalance vibration generator",
+        "texts": [
+            " In addition to the stationary force, the test bearing can be loaded with sinusoidal forces. For this purpose, vibration generators (10) are attached to the frame of the test bearing. During the test procedure the relative movements between shaft and bearing are measured in order to determine the dynamic behavior of the test bearing. 4.1.1 Dynamic forces In addition to the stationary force, the test bearing can be loaded with sinusoidal forces. For this purpose vibration generators (10 in Figure 4) are attached to the frame of the test bearing in \u00b145\u00b0 to the vertical. Figure 5 shows the mechanical design of the vibrator in a section view. It consists of two pairs of imbalanced shafts (2 and 3), which are supported in a massive housing (1). These are connected with a toothed belt drive, so that looking in direction of the flange (12), the two adjacent shafts rotate in opposite directions. The shafts of each pair are precisely in phase to each other in such a way that the resulting force component of the vibrator, transverse to the radial bearing direction is zero. Therefore the resulting force acts in radial bearing direction only"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001954_13506501jet521-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001954_13506501jet521-Figure5-1.png",
+        "caption": "Fig. 5 Temperature of feeding and oil film of two different working conditions ((a) seawater 3 per cent by mass; (b) rotational speed 1000 r/min; (c) rotational speed 2750 r/min, 1 per cent seawater content)",
+        "texts": [
+            " The stress and pressure of the lubricant were also predicted from the CFD code. The predicted frictional shear stress \u03c4 was integrated over the bearing area Sb to determine the friction force Ff of the oil film Ff = \u222b \u222b Sb \u03c4 dSb (1) The load capacity Fp was obtained by integrating the vertical component of the pressure pz over the journal area Sj Fp = \u222b \u222b Sj pz dSj (2) Figures 5(a) to (c) show the temperature of feeding and oil film for different rotational speeds, sea water contents, and running time periods. Figure 5(a) shows the effect of two different rotational speeds of 2750 and 500 r/min for a fixed 3 per cent sea water content, while the effect of the sea water content is presented in Fig. 5(b) for a fixed rotational speed of 1000 r/min. Figure 5(c) shows the effect of running time periods for a fixed rotational speed of 2750 r/min and 1 per cent sea water content. The feeding temperature was similar between the two rotational speeds of 2750 and 500 r/min. However, the oil film temperature at 2750 r/min was much higher than that at 500 r/min as shown in Fig. 5(a). The temperature rise increased along the rise of rotational speed, and this was in accordance with the theory. The feeding temperature for the 3 per cent sea water content was slightly lower than that for the 1 per cent sea water content, but the circumferential temperature of the oil film was similar between the two conditions. Therefore, the temperature rise for the 3 per cent sea water content was slightly higher than that for the 1 per cent sea water content. The elevated temperature was a result of a higher friction force due to the increased viscosity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003920_s11044-012-9320-0-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003920_s11044-012-9320-0-Figure9-1.png",
+        "caption": "Fig. 9 Sketch of the combined tilting and precessing hub case",
+        "texts": [
+            " This allows to eliminate the portion of \u03c9b associated to the tilting of body b, but not the portion associated with precession. As a consequence, when body b is subjected to pure tilting, \u03c9b must be equal to \u03c9a ; when body b also precesses with velocity \u03c9p , the expected velocity is \u03c9b = \u03c9a + (1\u2212 cos\u03b2)\u03c9p , where \u03b2 is the tilting angle. Consider an arbitrary combined tilting and precession motion consisting of forcing eb 3 about a conical trajectory where the axis of the cone is tilted by \u03b20 about ea 3 and the cone angle is \u03b21. The cone is spanned by eb 3 at constant angular velocity \u03c9t about the axis of the cone. Figure 9 illustrates the prescribed motion of axis eb 3. The closed form solution, eb 3 = (sin\u03b20 cos\u03b21 + cos\u03b20 sin\u03b21 cos\u03c81)e1 + sin\u03b21 sin\u03c81e2 + (cos\u03b20 cos\u03b21 \u2212 sin\u03b20 sin\u03b21 cos\u03c81)e3, (46) is given in [3], where \u03c81 = \u03c91t is the azimuth of the plane that contains ea 3 and eb 3. Figure 10 presents the numerical results. They match the analytical solution and the numerical solution of [3] to the desired level of precision. Note that in Fig. 10 \u03c9b is not constant since it necessarily contains a variable fraction of the precession velocity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.22-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.22-1.png",
+        "caption": "Fig. 3.22. Representation of linear and angular velocities in different coordinate frames on the same rigid body",
+        "texts": [
+            " If this manipulator is allowed to move, e.g., in the case Ke /\u2208 N (JT ), the end-effector attains the desired posture and the corresponding joint variables are determined. The kineto-statics duality concept presented above can be useful to characterize the transformation of velocities and forces between two coordinate frames. Consider a reference coordinate frame O0\u2013x0y0z0 and a rigid body moving with respect to such a frame. Then let O1\u2013x1y1z1 and O2\u2013x2y2z2 be two coordinate frames attached to the body (Fig. 3.22). The relationships between translational and rotational velocities of the two frames with respect to the reference frame are given by \u03c92 = \u03c91 p\u03072 = p\u03071 + \u03c91 \u00d7 r12. By exploiting the skew-symmetric operator S(\u00b7) in (3.9), the above relations can be compactly written as[ p\u03072 \u03c92 ] = [ I \u2212S(r12) O I ] [ p\u03071 \u03c91 ] . (3.112) All vectors in (3.112) are meant to be referred to the reference frame O0\u2013 x0y0z0. On the other hand, if vectors are referred to their own frames, it is r12 = R1r 1 12 and also p\u03071 = R1p\u0307 1 1 p\u03072 = R2p\u0307 2 2 = R1R 1 2p\u0307 2 2 \u03c91 = R1\u03c9 1 1 \u03c92 = R2\u03c9 2 2 = R1R 1 2\u03c9 2 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003060_s0025654411050141-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003060_s0025654411050141-Figure7-1.png",
+        "caption": "Fig. 7.",
+        "texts": [
+            " Figure 4 shows the time dependence of the body displacement (deformation) and the contact force of interaction between the body and the obstacle for V \u2212 = 1 m/s, 2 m/s, 3 m/s, 4 m/s, and 5 m/s. Figure 5 displays the dependence of the body maximum displacement (maximum deformation), maximum contact force, and the impact duration on the collision velocity V \u2212. Figure 6 presents the dependence of the contact force on the body displacement (deformation) for V \u2212 = 1 m/s, 2 m/s, 3 m/s, 4 m/s, and 5 m/s. The area bounded by this curve is equal to the loss of the kinetic energy at the impact. The dependence of the kinetic energy lost at the impact on the collision velocity V \u2212 is shown in Fig. 7. MECHANICS OF SOLIDS Vol. 46 No. 5 2011 LAMBERT W -FUNCTION This function became widely known after it was introduced in the computer algebra system MAPLE in the 1980s [6]. Earlier, it was used by some authors, beginning from Euler, to solve different mathematical problems. This function is differentiable and integrable, and there are effective procedures for its computer implementation. The function W has been used to obtain analytical solutions for a number of mathematical and mechanical problems [6]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002986_s00170-012-4659-1-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002986_s00170-012-4659-1-Figure1-1.png",
+        "caption": "Fig. 1 Schematic of SCMW",
+        "texts": [
+            "eywords Electrochemical brushing . Space curve meshing wheel . Instantaneous transmission ratio . Precision meshing transmission The space curve meshing wheel (SCMW) is a novel gear mechanism designed according to the space curve meshing theory [1\u20134]. It could perform successive meshing transmission with two crossed shafts. An SCMW transmission model was established, as shown in Fig. 1. The transmission ratio of this model is 4. The angle of the angular velocity vectors of the driving wheel and driven wheel is 150\u00b0. These SCMW samples were manufactured by the selective laser melting (SLM) rapid prototyping process [5]. The instantaneous transmission ratio was measured by the kinematics experimental rig and the result is illustrated in Fig. 2 [1].The SCMW performs transmission by the meshing tines symmetrically disposed on the driving wheel and driven wheel. The surface of the meshing tines was rough (Ra034 \u03bcm)",
+            " Then, the ECB process was used to finish these Ingredients of the electrolyte (%) NaNO3 (10 %)+H2O+ Al2O3 (0.1 %) Material of the cathode tool 316L stainless steel Material of the workpiece 316L stainless steel Machining gap 1 mm Machining voltage 10 V Roughness of the workpiece surface 34 \u03bcm Roughness of the cathode tool surface 0.3 \u03bcm SCMW samples and the main machining parameters are shown in Table 1. Cathode tools were designed according to the shape of the SCMW samples, the relative motion of the cathode tool, and the workpiece. 3.1 Finishing process for driving tines As shown in Fig. 1, the central line of the driving tine of the SCMW samples is a cylindrical helix. And the equation of the central line is expressed in Eq. 3 [1]. The section of the driving tine is circle. Suppose the diameter is d: x1 \u00bc m cos t y1 \u00bc m sin t z1 \u00bc nt \u00fe np p t p 2 8< : \u00f03\u00de wherem is the radius of the spiral curve and 2n\u03c0 is the pitch of the spiral curve. The schematic diagram of the ECB finishing system designed for the driving tines is shown in Fig. 11. As shown in Fig. 12, the machining tool has six machining areas which match the driving tines"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000762_tsmcb.2007.909943-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000762_tsmcb.2007.909943-Figure2-1.png",
+        "caption": "Fig. 2. Rough plate and plane sphere.",
+        "texts": [
+            " For ideal geometries, the van der Waals forces are given by FVdW bp = Aw bpRb 6D2 bp FVdW bs = Aw bsRb 6D2 bs FVdW bb = Aw bbRb 6D2 bb (4) for ball\u2013probe (bp), ball\u2013substrate (bs), and ball\u2013ball interactions, respectively. Here, Rb is the sphere radius, Aw ij is the Hamaker constant of the \u201ci\u2212water\u2212j\u201d interface, and Dij is the separation distance. Furthermore, van der Waals forces are greatly influenced by the surface roughness [2]. It has been shown that increasing the surface roughness decreases the van der Waals forces [4]. Thus, taking the surface roughness into consideration, as shown in Fig. 2, the van der Waals force is expressed as [6] Fvdwb = ( z z + b/2 )2 Fvdw (5) where z is the distance, b is the height of the surface irregularities, and Fvdw is the van der Waals forces between the plane plate and the sphere. In ambient operational environment, the water layer is present on the surface of the sphere and the substrate. A liquid bridge occurs between them at close contact, as shown in Fig. 3. In [20], the macroscopic theory of capillarity is proven to be applicable for a curvature radius on the order of molecular size"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002997_detc2011-48348-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002997_detc2011-48348-Figure4-1.png",
+        "caption": "Figure 4. Interaction of human factors considerations [70]",
+        "texts": [
+            " There is a need for clearer communication in order to avoid this kind of disputes and further delays in time-to-market. The importance of communication is also addressed by Ernst [67], who identified intensive communication in the design team as a success factor for new product development (NPD). The involvement of the user in the development has been defined as a success factors for NPD [67] and for the specific case of medical devices [68, 69]. In addition to their involvement, a variety of human factors aspects should be considered during the validation and verification of medical devices. Figure 4 illustrates FDA\u201fs representation of the human factors considerations to evaluate along with the device use, which include the environment, the user and the device. The process recommended includes the identification of use-related hazards and their mitigation through the modification of device user interfaces or the abilities of users to use the device [70]. Note that hazards considered in the risk analysis refer more to the device failure hazards, which may include chemical, mechanical thermal, electrical, radiation or biological hazards"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003578_j.cad.2011.03.001-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003578_j.cad.2011.03.001-Figure3-1.png",
+        "caption": "Fig. 3. The inscribed disks and the medial axis disks (bold line).",
+        "texts": [
+            " As discussed earlier, the radius of themedial axis disk comes from the value of the distance function, so the MAT has direct relation with the extremal property of the distance function. Now considering the above condition (a), let p(u, v) be a point in \u2126 , we use the notation r(u, v) to represent the minimum of the distance function r(u, v) = min t |q(t) \u2212 p(u, v)| . (3) From Eq. (3), we know that the inscribed disks with radii r that are the minimum distances are wholly contained in \u2126 , as shown in Fig. 3. Suppose p is a moving point in \u2126 , among a family of inscribed disks centered at p, the largest one must possess the maximum property according to the condition (b), the value of the distance function can be expressed as rs = max u,v min t |q(t) \u2212 p(u, v)| . (4) From Eq. (4), we know that the distance function is minimumwith respect to t , andmaximumwith respect to u and v. Considering the minimum\u2013maximum property possessed by the above distance function, the inscribed disk must be the largest one in domain \u2126 . As shown in Fig. 3, the inscribed disk with radius rs centered at point ps is the largest one, and the center ps is a saddle point of the distance function. Since the coordinate parameters u and v of Eq. (4) are two independent variables, the point p can move within \u2126 with two degrees of freedom during the searching process for themaximum r . If we impose some restrictions on u or v, such as v = v(u) or u = u0 (constant) or v = v0 (constant), and in this case we define the point to be conditional saddle point. Here, only u = u0 is taken as an example, Eq. (4) can be rewritten as rcs = max v min t |q(t) \u2212 p(u0, v)| (5) where p has only one degree of freedom, i.e., the center of the inscribed disk is only allowed to move in direction v. In Fig. 3, the center pcs of the maximal inscribed disk which is bitangent to the boundary is a conditional saddle point. It is important to remark that the notations \u2018\u2018max\u2019\u2019 and \u2018\u2018min\u2019\u2019 in Eqs. (4) and (5) are global maximum and global minimum. Using them simultaneously to a distance function means that the minimum distance function has been maximized globally. That is the basic property of the saddle points and the conditional saddle points. In addition, there is another saddle point called local saddle point, where the globalminimumof the distance function is maximized locally. See Fig. 3, the center pls is a local saddle point, the corresponding radius rls is the global minimum but maximized locally. Generally, there is only one saddle point but there aremany local saddle points in a closed domain. From the above discussion, we know that the deeper meaning of the MAT can be highlighted by using saddle point, local saddle point and conditional saddle point. The conditional saddle points of the distance function correspond to the regular points of the MA; the saddle points and local saddle points correspond to the branch points; the radius of medial axis disk equals to the value of the distance function. All these can be called the saddle point properties of theMAT. In Fig. 3 the conditional point pcs is a regular point of the MA, the saddle point ps and local saddle point pls are branch points, the disks (bold line) are the corresponding medial axis disks. As shown in Fig. 4, suppose p0 is an initial medial axis point, and q1, q2 are two associated tangent points between the medial axis disk and the boundaries C1 and C2, respectively. Frames {p0, e1e2}{q1, e11e 1 2} and {q2, e21e 2 2} are Frenet frames of MA, C1 and C2, with their origins at points p0, q1 and q2, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000021_esda2006-95565-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000021_esda2006-95565-Figure6-1.png",
+        "caption": "FIGURE 6. GUTTER BRUSH OPERATING PARAMETERS",
+        "texts": [
+            " In contrast, a flicking brush bristle (\u03b3 = 90\u00b0) mainly deflects in the tangential direction, \u201cflicking\u201d debris when it is released from the road at relatively high speed. 3 Copyright \u00a9 2006 by ASME rms of Use: http://www.asme.org/about-asme/terms-of-use Do Apart from the material properties and geometric parameters, the performance of the brush depends on the operating parameters. Some of them are the brush offset angle, \u03be, brush angle of attack, \u03b2, brush rotational speed, \u03c9, vehicle speed, v, brush penetration, \u2206, brush vertical force, Fb, and brush torque, T, which are illustrated in Fig. 6. The mechanics of cutting and/or flicking brushes have been studied by Vanegas Useche et al. [6,16], Peel and Parker [8], Peel et al. [14], and Peel [15]. As many of the results on this are unavailable as conference or journal papers, it is important to discuss them here to a certain extent. It is noted that the results of [6] are reviewed by the authors in [16], and, thus, they are not presented here. Operating parameters such as forces and torques in a horizontal or tilted brush that acts against flat surfaces (oil lubricated steel and motorway grade concrete) without debris are analysed through a static model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001722_iet-gtd.2009.0015-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001722_iet-gtd.2009.0015-Figure6-1.png",
+        "caption": "Figure 6 Superimposed component net of the transmission line with shunt reactor when external fault occurs",
+        "texts": [
+            "1 Analysis of the shunt reactor influence As to actual UHV/EHV lines, shunt reactors are usually used to compensate capacity currents and prevent over voltage, so it is necessary to analyse the shunt reactor influence. The internal fault and external fault of transmission lines with shunt reactor are shown in Figs. 5 and 6, respectively, in which L is the inductance of shunt reactor, and other symbols are the same as ones in Figs. 1 and 3. According to Fig. 5, (10) is still correct, thus the shunt reactor has no influence in this new approach during internal fault, the protection will still operate correctly. On the other hand, according to Fig. 6, the situation is quite different. For terminal m Dim Di0m \u00bc C dDum dt \u00fe \u00f0t 0 Dum(t) L dt (15) For terminal n Din Di0n \u00bc C dDun dt \u00fe \u00f0t 0 Dun(t) L dt (16) Since Di0m \u00fe Di0n \u00bc 0, the following equation is obtained Dicd \u00bc C dDucd dt \u00fe \u00f0t 0 Ducd (t) L dt (17) 8 The Institution of Engineering and Technology 2009 Comparing (4) with (17), it is obvious that they are not identical. In this EMTP simulation, the superimposed component differential currents with inductance compensation, Di00m and Di00n , are used to compute protection criterion instead of Dim and Din, considering that the inductance compensation current is usually available in practice, where Di00m \u00bc Dim \u00f0t 0 Dum(t) L dt, Di00n \u00bc Din \u00f0t 0 Dun(t) L dt and Dicd \u00bc Di00m \u00fe Di00n In order to verify the effectiveness and reliability of the scheme proposed, one 400-km long 500 kV transmission line system with shunt reactor of 70% compensation degree is employed as simulation model as illustrated in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003530_amr.694-697.503-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003530_amr.694-697.503-Figure1-1.png",
+        "caption": "Fig. 1 Basic normal tooth profile of involute",
+        "texts": [
+            " According to the meshing theory, the spiral bevel gear for nutation drive can be seen as the enveloping process by an imaginary rotating crown gear with respect to the bevel gear [7]. The mathematical model of involute spiral bevel gears would be obtained by substituting the equation of involute basic normal tooth profile into the general mathematical model of internal meshing spiral bevel gears. The complete involute basic normal tooth profile is composed of the involute curve and the dedendum transition curve, as shown in Fig. 1. Coordinate system ),,( zyx is built on the base circle center of involute curve, and the coordinate system ),,( nnn zyx establish on the intersection point of All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 141.211.4.224, University of Michigan Library, Media Union Library, Ann Arbor, USA-08/07/15,16:25:27) base circle and x axes. According to the forming principle of involute, the basic parameters have been given as: pressure angle \u00b0= 20\u03b1 , addendum coefficient 1=\u2217 ah , tip clearance coefficient 25"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002494_1559-0410.1299-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002494_1559-0410.1299-Figure1-1.png",
+        "caption": "Figure 1. Dimensions",
+        "texts": [
+            " Their model has been tested extensively, producing reliable results with errors in basketball simulations of less than 1%, and is used throughout this paper. The dimensions of the court, backboard, and ring that influence the bank shot are the same for international competition (International Basketball Federation, 2006), US collegiate competition (National Collegiate Athletic Association, 2001), and US professional competition (National Basketball Association, 2006). However, the conclusions reached in the present study apply only to men\u2019s basketball because in woman\u2019s basketball the ball is smaller and lighter (Fig. 1). 2 DOI: 10.2202/1559-0410.1299 Brought to you by | Dalhousie University Authenticated Download Date | 5/17/15 4:48 PM In Figure 2, the ball is launched from a particular location that can be expressed in terms of the rectangular coordinates (x y z) or equivalently in terms of the cylindrical coordinates (r  z) in which r denotes radial distance and  denotes polar angle. The coordinates are located at the center of the ball. The ball is launched with a launch speed v, a launch angle  and an aim angle "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002063_robio.2010.5723474-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002063_robio.2010.5723474-Figure1-1.png",
+        "caption": "Fig. 1. 7DOF Redundant Manipulator",
+        "texts": [
+            " To avoid this, this paper deduces a motion planning algorithm of the manipulator. This paper will be organized as below. The basic description of this 7DOF redundant manipulator and the forward kinematic model will be introduced in Section II. Secondly, the online motion planning algorithm will be deduced in Section III. In Section IV, a simulation result will be given. Finally, the conclusion and future work will be demonstrated in Section V. II. FORWARD KINEMATIC OF 7DOF REDUNDANT MANIPULATOR The 7DOF redundant manipulator is shown in Fig. 1. We can see that all the joints of this manipulator is revolve joint and any two axis of each joint is perpendicular to each other as well as cross in one point. Besides, the gripper is attached to the 7th link. To build the forward kinematic model, the coordinate of every link is defined base on a classical Denavit-Hartenberg model [7]. The base coordinate frame of the manipulator is defined as \u03a30. Then, each joint i(i = 1, 2...6) is defined as \u03a3i based on the Denavit-Hartenberg rules (DH rules)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001386_cdc.2009.5400224-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001386_cdc.2009.5400224-Figure3-1.png",
+        "caption": "Fig. 3: Switching surface",
+        "texts": [
+            "27) implies that there exists a constant \u03b5 > 0 such that \u03bb\u0307i(t) < \u2212\u03b5 i f \u03bbi(t) > 0 \u03bb\u0307i(t) > \u03b5 i f \u03bbi(t) < 0. (4.30) Therefore, there exists a time t\u2217 > 0 such that \u03bbi(t) = 0 \u2200t \u2265 t\u2217. (4.31) Notice that the closed-loop system (2.1), (4.25) is a system of differential equations with discontinuous right-hand sides, the equation \u03bbi = 0 defines a switching surface of this system, a solution satisfying (4.31) is a sliding mode (see e.g. [14]). The inequalities (4.30) guarantee that the vector field of the closed-loop system around this switching surface looks as it is shown in Fig. 3. Furthermore, any solution satisfying (4.31) satisfies \u03bb\u0307i(t) = 0 \u2200i \u2200t \u2265 t\u2217. (4.32) From this and (4.26), we obtain that \u03c9i(t) = \u2212 sin(\u03b2i j(t))\u2016V T i j (t)\u2016 \u2016di j(t)\u2016 (4.33) for all sliding mode solutions. Therefore, for any initial condition, the sliding mode solution is unique and welldefined. Furthermore, (4.33), (4.29), (4.19) and (4.27) imply that the constraint (2.2) holds for any sliding mode solution satisfying (4.31). Furthermore, the condition (4.31) means that the velocity vector Vi(t) is parallel to the vector di(t) for all t \u2265 t\u2217"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003137_6.2010-3417-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003137_6.2010-3417-Figure8-1.png",
+        "caption": "Figure 8. Adaptive control effort for the standard adaptive law with \u0393std = 1x10\u22122 and \u0393std = 1x10\u22127.",
+        "texts": [],
+        "surrounding_texts": [
+            "When fast adaptation is desired for systems with slow reference models or slow dynamics, undesirable high magnitude oscillatory control signals can result. A neural network adaptive control architecture is presented that allows the possibility of fast adaptation without oscillatory control signals and with smooth weight convergence. Simulation results using a model for wing rock and a Boeing 747 model illustrate the methods power. Future work will include expanding this methodology to band limit the entire adaptive control signal within the H\u221e Adaptive Control framework17 ."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000923_bfb0110378-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000923_bfb0110378-Figure6-1.png",
+        "caption": "Fig. 6. n-trailer system (left) and 1-trailer system with kingpin hitch (right).",
+        "texts": [
+            " Since it is fiat when a = 1 it is orbitaIly fiat when a 7 s 1, with (qD, ~) as an orbitally fiat output. Such systems are fiat by theorem 2 since they correspond to dr@less systems with n states and n - 2 inputs. For instance the rolling disc (i9. 4), the rolling sphere (p. 96) and the bicycle (p. 330) considered in the classical treatise on nonholonomic mechanics [51] are flat. Example 17 (Mobi le robo t s ) Many mobile robots modeled by rolling without sliding constraints, such as those considered in [5, 50, 76] are fiat. In particular, the n-trailer system (figure 6) has for fiat output the mid-point Pn of the last trailer axle [68,12]. The 1-trailer system with kingpin hitch is also fiat, with a rather complicated fiat output involving elliptic integrals [67,10], but by theorem 1 the system is not flat when there is mote than one trailer. Example 18 (The rolling penny) The dynamics of this Lagrangian system submitted to a nonholonomic constraint is described by :~ = A sin p + ul cos = -A cos~ + ul sin~ : U 2 xsinqo = ~) cos~ where x, y, qo are the configuration variables, A is the Lagrange multiplier of the constraint and ux, u2 are the control inputs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003775_elan.201200574-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003775_elan.201200574-Figure1-1.png",
+        "caption": "Fig. 1. Manufacturing methods for electrode miniaturization: a) ink-mask deposition; b) metallization and heat treatment for improving adherence of the metal film; c) ink-mask removal; d) deposition of the reference electrode; and e) final array structure.",
+        "texts": [
+            " The electrochemical response of the films was also obtained by using a Stanford QCM200 microbalance (EQCM) with a three-compartment electrochemical Teflon cell, Au-QCM as the working electrode, Ag/AgCl as the reference electrode, and a Pt plate (2.0 cm2) as the counter electrode. The working electrodes used here were 5 MHz AT-cut quartz crystals coated with polished Au films on both sides (Au-QCM); and also, Au microelectrodes built on polyamide substrates and prepared by lithography using a laser printing. The Au microelectrodes were built on polyamide substrates, which were previously prepared by lithography with a laser printed mask (Figure 1 a). The Au layers were evaporated onto the substrate surfaces and then, heated at 240 8C during 10 min in order to improve adhesion (Figure 1 b). The printing ink was dissolved in acetone, and then, rinsed, leaving the Au tracks totally clean (Figure 1 c). After obtaining the Au tracks, the electrode area was delimited by using epoxy resin and previously deposited Ag layers, which were electrochemically prepared in an aqueous 0.1 molL 1AgNO3 and AgCl in 0.1 molL 1 HCl solution (Figure 1 d). As a result, an electrochemical cell was obtained, including the Au working electrode with a radius of 500 mm, Au counter electrode, and Ag/AgCl reference electrode (Figure 1 e). The potential of the Ag/AgCl microelectrode in 0.05 molL 1 phosphate buffer at pH 6.5 was measured as 87 9 mV vs. Ag/ AgCl. All electrochemical measurements made by using the biosensor arrays were performed with a flow injection cell with a total volume of 100 mL built with two rubber o-rings between two Teflon plates separated by a microelectrode (Au). When inside the electrochemical cell, the Teflon plate was fixed facing on the Au surface, which allowed both inlet and outlet solution fluxes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003199_0954410011403578-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003199_0954410011403578-Figure1-1.png",
+        "caption": "Fig. 1 Canard aerodynamic model force diagram",
+        "texts": [
+            " These equations are well known and reported in many sources [6] _x _y _z 8< : 9= ; \u00bc c c s s c c s c s c \u00fe s s c s s s s \u00fe c c c s s s c s s c c c 2 4 3 5 u v w 8< : 9= ; \u00f01\u00de _ _ _ 8< : 9= ; \u00bc 1 s t c t 0 c s 0 s =c c =c 2 4 3 5 p q r 8< : 9= ; \u00f02\u00de _u _v _w 8< : 9= ; \u00bc X =m Y =m Z=m 8< : 9= ; 0 r q r 0 p q p 0 2 4 3 5 u v w 8< : 9= ; \u00f03\u00de _p _q _r 8>< >: 9>= >; \u00bc Ixx Ixy Ixz Ixy Iyy Iyz Ixz Iyz Izz 2 64 3 75 1 L M N 8>< >: 9>= >; 2 64 0 r q r 0 p q p 0 2 64 3 75 Ixx Ixy Ixz Ixy Iyy Iyz Ixz Iyz Izz 2 64 3 75 p q r 8>< >: 9>= >; 3 75 \u00f04\u00de The applied forces and moments in equations (3) and (4) contain contributions from the weight of the projectile, body aerodynamic forces, and canards [5]. The aerodynamic force due to a single canard is modelled as a point force acting at the lifting surface aerodynamic centre, as shown in Fig. 1. The orientation of a particular canard is obtained by one bodyfixed rotation. Starting with the canard axis aligned with the projectile body axis, the canard is rotated about the ~IB axis by the azimuthal \u00f0 Ci \u00de angle. Figure 2 shows a diagram of the two canards used in this development. Strip theory is used to compute the canard aerodynamic loads [6]. Notice in Fig. 1 that the aerodynamic angle of attack of the ith canard is calculated using only the uACi and wACi components of the relative air velocity experienced by the canard computation point. In the projectile body axis, the ith canard force is given by equation (5). The transformation matrix TCi is shown in equation (6). XCi YCi ZCi 8>< >: 9>= >; \u00bc qCi SCi TCi CLCi sin\u00f0 Ci i\u00de CDCi cos\u00f0 Ci i\u00de 0 CLCi cos\u00f0 Ci i\u00de CDCi sin\u00f0 Ci i\u00de 8>< >: 9>= >; \u00f05\u00de TCi \u00bc 1 0 0 0 cos\u00f0 Ci \u00de sin\u00f0 Ci \u00de 0 sin\u00f0 Ci \u00de cos\u00f0 Ci \u00de 2 4 3 5 \u00f06\u00de Proc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003014_iros.2011.6094585-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003014_iros.2011.6094585-Figure4-1.png",
+        "caption": "Fig. 4. Assembly of Active Split Offset Caster module.",
+        "texts": [
+            " The onboard computer supervises all ASOC modules via Xbee wireless links, such that the onboard computer kinematically coordinates ASOC modules to move the robot in a desired direction. GPS data is collected as ground truth data for outdoor experiments. These data are sent to an operator via IEEE 802.11g, along with ASOC motion data (i.e. potentiometer for pivot angle and wheel tachometer data). B. Active Split Offset Caster Module Description The ASOC module consists of a split wheel pair, connecting axle, and offset link connected to the wheel pair (Fig. 4). Based on the kinematic isotropy analysis reported in [13][14], the geometric ratio offsetsplit LL is designed to be 2.0 for the most isotropic mobility the ASOC. The wheel pair/axle assembly passively rotates around the pivot axis. The roll axis also passively rotates, maintaining wheel contact on sloped or rough terrain surfaces. The angle of rotation of the pivot and roll axes are measured by potentiometers. produce planar translational velocities at a point along its pivot axis by independently controlling each wheel\u2019s velocity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure3-1.png",
+        "caption": "Figure 3. Rolling element thrust bearings",
+        "texts": [
+            " The gear contact analysis includes a consideration of the calculated gear mesh misalignment, a non-linear tooth stiffness, as well as the detailed micro-geometry surface definition. The predicted position of the contact on the tooth flank will influence the effect of the gear mesh force. For example, if the contact is away from the centre of the tooth, this will generate a turning moment on the gear, which will have to be reacted by the supporting bearings. The rear planetary gear set in this example is shown in Figure 2. Within the model, the axial load is properly taken by the addition of rolling element thrust bearings (Figure 3), as shown in Figure 4. The clutch and brake parts within the system are modeled as mass, inertia and stiffness components. The completed model is shown in Figure 6, including all gears, bearings, planet carriers and clutches. The boundary conditions of the analysis are then defined. This entails specifying the input speed and torque (or power), and which clutches or brakes are locked. The convergence scheme is shown in Figure 7. The non-linear analysis can be completed in a couple of minutes. The results show the full six degree of freedom deflection of the system, the rotational speeds of and forces acting on all components, gear and bearing misalignments"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure1-1.png",
+        "caption": "Figure 1. Parameters included in the modelling of a taper roller element bearing.",
+        "texts": [
+            " The modelling time for those components and assemblies are significantly reduced and possible mistakes, which may inevitably happen in a conventional modelling process, may be avoided. The engineer does not have to consider details such as defining the powerflow for a particular speed, the direction of gear mesh forces, or the stiffness of non-linear loaded bearings. Rolling element bearings are added to the model by selecting the appropriate component from a pre-defined catalogue. The model includes a detailed representation of the rolling element internal details (as shown in Figure 1) and allows for the non-linear stiffness to be determined for any load condition, developed from standard methods [6]. The stiffness sub-matrix, linking the displacements and tilts of the inner and outer raceway geometric centres, is obtained as the slope of the force versus deflection curve at the bearing\u2019s operating displacements [7]. The stiffness terms include the contacts of the rolling elements with the raceways, the non-linear effects of internal clearance, and preload. The centrifugal effects in high-speed bearings are also considered"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002041_tac.1967.1098565-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002041_tac.1967.1098565-Figure9-1.png",
+        "caption": "Fig. 9. Portrait of optimal trajectories when singular controls are not possible.",
+        "texts": [
+            " The only possible control Since the control strategy which results in the optimal transition from any initial state x1 =xlO, x2 = x 2 0 to x1 =x2 = O must be the dual of the control strategy which results in the optimal transition from any initial state x1 = - x10, x2 = - x20 to x1 = x2 = 0, the trajectories shown in the region to the right of the curve A-kA+ in Fig. 7 are immediately verified. In other words, the geometrical curves defining the optimum-solution trajectories have odd symmetry with respect to the origin of the x1x2 plane. The above results establish the control policy shown in Fig. 8 when singular control is possible. IT,'hen singular control is not possible, it is easily verified by arguments identical to those employed above that the portrait of optimal trajectories is as shown in Fig. 9. The control policy in this case is shown in Fig. 10. The equation for the BTB'segment of the switching boundaries shown in Figs. 8 and 10 is readily derived when i t is noted that the switching boundary has odd symmetry and that the segment BT is determined by the system trajectory for zt = - L7 which passes through the origin. Taking the ratio of ( 3 ) and (4) with u = - li leads to dxp u + ax2 I x2 I - = - (34) dxl 2 2 or xzdxz. dzl = - u + ax2 i x2 I In the region x2>0, integration of (35) yields 1 2a x1 = - -1n (U + ax2) + C, where C is the constant of integration",
+            " When this peed is attainable, it is reasonable to presume that the best control policy under the constraint I u I 5 U is to achieve this velocity as rapidly as possible, then cruise, and finally decelerate as rapidly as possible to attain zero final velocity. This is precisely the motion called for by the trajectories in Fig. 7. iYhen the cruising speed xpS is not attainable, an optimal cruising speed is not possible. I t is logical to expect in this case that the optimal control policy is to minimize the translation time T. The \"bang-bang\" controller demanded by the trajectories of Fig. 9 is therefore not unexpected. COXLESIONS The problem of optimum rectilinear t anslation (against drag forces proportional to the square of the speed) is solved for the case of amplitude-constrained control inputs and performance index (1). The existence of a minimum value for the performance index is assumed in the solution and justified with a heuristic argument. APPESDIX From (1) and (2), i t follows that an equivalent expression for the performance index is 1 2 0 T S = - [ 9 ( T ) - k2(0)] + s (k + a 9 I i I ) d t "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001780_00368790910960066-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001780_00368790910960066-Figure1-1.png",
+        "caption": "Figure 1 (a) Non-recessed symmetric hole-entry type journal bearing system, (b) two-lobe non-recessed symmetric hole-entry type journal bearing system, and (c) enlarged view of rough bearing surfaces",
+        "texts": [
+            " 1 andfy\u00f0L h; 1=g\u00de \u00f02a\u00de fs \u00bc \u00f02 Vrj 2 1\u00deFs; since Vrj \u00fe Vrb \u00bc 1 \u00f02b\u00de where Fs is positive function of \u00f0L:h\u00de and surface pattern parameter g of a given surface and is expressed as: Fs \u00bc A1\u00f0L h\u00dea1 exp2a2\u00f0L h\u00de\u00fea3\u00f0L h\u00de2 forL h # 5:0 Fs \u00bc A2exp20:25\u00f0L h\u00de forL h . 5:0 \u00f02c\u00de where C, r, A1, A2, a1, a2 and a3 appearing in (2a)-(2c) are constants and g is the surface pattern parameter. The expression for non-dimensional fluid film thickness for a misaligned roughened multi-lobe rigid journal with reference to fixed co-ordinate axis shown in Figure 1(c) is given as (Ramesh et al., 1997, Nagaraju et al., 2006): hT \u00bc h 2 1 \u00fe erf L hffiffi 2 p \u00fe 1 L ffiffiffiffi 2p p e2\u00f0L h\u00de2=2 for L h , 3 h for L h $ 3 8< : \u00f03\u00de where L\u00f0\u00bc 1= s\u00de in the present work is defined as surface roughness parameter and h is the nominal fluid-film thickness for a misaligned smooth surface journal bearing system. The conditions \u00f0L h , 3\u00de and \u00f0L h $ 3\u00de characterize the partial and fully lubricated regions, respectively. The expression for non-dimensional nominal fluid film thickness Two-lobe hole-entry hybrid journal bearing J",
+            " Sharana Basavaraja, Sathish Sharma and Sathish Jain Volume 61 \u00b7 Number 4 \u00b7 2009 \u00b7 220\u2013227 D ow nl oa de d by U ni ve rs ity o f So ut he rn Q ue en sl an d A t 0 0: 03 2 7 Ju ne 2 01 6 (P T ) h in a misaligned multi-lobe hole-entry hybrid journal bearing system can be obtained be using transformation from bearing fixed axes (x, y, z) to journal fixed (x 0, y 0, z 0) and is expressed as (Ghosh and Satish, 2003, Singh et al., 1973): h \u00bc 1 d 2 \u00f0 Xj \u00fe x2 X i L\u00decosa2 \u00f0 Zj \u00fe z2 Z i L\u00desina \u00fe cb cosa2 wb sina \u00f04\u00de where Xj and Zj are the equilibrium co-ordinates of the journal center, and x and y are time dependent perturbation co-ordinates of the journal center measured from its equilibrium position. X i L and Z i L are the lobe center co-ordinates of ith lobe (Figure 1). The flow rate of the lubricant through the capillary restrictor is defined as (Sharma et al., 2001): QR \u00bc CS2\u00f01 2 pc\u00de \u00f05\u00de where the term pc represents the pressure at the hole. The lubricant flow field in the clearance space of a circular bearing has been discretised into four nodded isoperimetric elements and using the Lagrangian interpolation function, the pressure at a point in the element is bilinearly distributed and expressed approximately as (Hubner, 1975): p \u00bc X4 j\u00bc1 Nj pj \u00f06\u00de using the approximate value of p in equation (1) and applying Glalerkins techniquewithusual assemblyprocedure, the following global system equation is derived as (Sharma et al"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002841_j.tsf.2012.04.054-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002841_j.tsf.2012.04.054-Figure1-1.png",
+        "caption": "Fig. 1. Schematic drawing of the formation process of a circular silicone oil substrate. The inner circle represents the mark of the original silicone oil drop dipped on the glass surface before deposition. The outside circle represents the spreading front of the silicone oil during deposition.",
+        "texts": [
+            " The samples were prepared by using the direct current magnetron sputtering method at room temperature. A drop of silicone oil (DOW CORNING 705 Diffusion Pump Fluid with a vapor pressure below 10\u22128 Pa at room temperature) with a diameter 1\u20132 mm was dripped on a clean glass surface before deposition. The key phenomenon used in our experiment is that the oil drop can spread outward at the early stage of film deposition due to the bombardment of depositing atoms and the heating of target source [19,20]. The spreading process is sketched in Fig. 1. About 10 min later, a continuous Co film forms on the liquid surface and the spreading of the oil drop stops. The diameter of the resulting oil substrate is 4\u20136 times larger than that of the original oil drop. The sputtering target was a cobalt disk (purity 99.9%) with diameter D=60 mm. The target\u2013substrate distance was fixed to be 90 mm. The residual gas pressure before sputtering was about 4\u00d710\u22124 Pa. The films were deposited under argon (purity 99.999%) pressure of 0.5 Pa. The sputtering power P was fixed to be 50 W and 100 W"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001447_robot.2007.363852-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001447_robot.2007.363852-Figure1-1.png",
+        "caption": "Fig. 1. Suture Model and Virtual Coupling of Haptic Device",
+        "texts": [
+            " Further more, there is no discussion about collision detection and force propagation for haptic interaction between the user and the suture model. Our suture model is built based on all the forces mentioned in [12], and we provide a user-interface to allow users tie an arbitrary knot. Also we analyze how the forces propagate along the suture during knotting and unknotting. With the virtual coupling technique[13], we can provide very smooth force feedback to the user. For ID element, we model our suture as a mass-spring system which consists of a sequence of mass points laying on the centreline of the suture. (see (a) of Fig. 1). During graphic rendering, we use cylinders as suture segments connecting two successive points. We use Euler method to calculate the shape of our suture. We first compute the total force acting at each point, Pi, and then update its position based on the computed force. Once the total force at each of the nodes has been calculated, with the interval time dt, we can obtain the velocity and position of each point. The following part of this section explains the forces we simulate in our simulator",
+            " However, a position controlled impedance style haptic device, such as PHANTOM Omni and PHANTOM Desktop from Sensable Tchenology, forces are not directly available as input variables into the model. Furthermore, the mechanical characterization and digital nature of the haptic device make the operation of directly incorporating the device as part of the simulation more challenging. To overcome these difficulties, we use virtual coupling technique which introduces a indirect layer of interaction between the mechanical device and the simulation by employing a spring-damper between a simulated body and the device end-effector (see (b) of Fig. 1). Another advantage of this technique is that we can use different constants for computing the output force for the device versus the input force for the simulated body, which makes the forces appropriate for both the haptic device and the dynamic simulation. 3) Friction Force: In this paper, we only consider coulomb and viscous friction forces during the procedure of knotting and unknotting, and we do not consider the rolling friction. We will study the static friction and focus on how to tie a knot tightly and unknot a tight knot in the future"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000828_978-3-642-65590-6_3-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000828_978-3-642-65590-6_3-Figure2-1.png",
+        "caption": "Fig. 2 shows a short segment BEe of the centreline of an elongated animal supposed to be performing undulatory motions from side to side in a stream of velocity U in which those swimming motions just stop it slipping bacJrwards. (Thus U is its mean swimming speed relative to the stream). The undulation passes backwards at a different speed V. The lateral velocity is W so that in time r the centre-line moves sideways a distance Wr from A to B, while the wave moves backwards a distance Vr from A to C and therefore C is at the lateral position occupied by A a time r earlier. During the time interval r the water moves a distance Ur so that relative to the water the centre-line moves a distance DE = wr, where w, its normal velo city relative to the water, is",
+        "texts": [
+            " Here I shall take just a few minutes to expound them in by far their simplest form when the resi&tance relationships (whatever the Reynolds number) are approxi mated by linear laws and the cosine of the angle between the animal's line of centres and its direction of motion is approximated by-I. This conclusion is as in Eq. (2) and Fig. 5 of Ligktkill (1969); note also an alternative, analytical way of appreciating it using k(x, t) for lateral displacement as in Ligktkill (1960) and the equa tions W = ok/at, w = ok/Of + U ok/ox, 0 = ok/at + v ok/ox. (2) We take force per unit length of centre-line resisting this normal velocity was KNw, where KN is a normal-force coefficient, and multiply that by the slope WjV (see Fig. 2) to obtain the corresponding component of forward tkrust per unit length K NW( W / V). The same slope means however that the lateral velocity W includes a tangential component W(WjV) which with KT as the tangential-force coefficient reduces the thrust per unit length by KTW(WjV) to the net amount (3) This thrust per unit length (3) averaged along the whole centreline must balance the drag per unit length KTU. Equation (3) is a formula of which two uses can be made. The useful thrusting work UP done per unit length can be compared with the rate W(KNW) of mechanical energy expenditure by the lateral motions with velocity W against the resisting normal force KNw per unit length"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002447_s2238-7854(12)70004-2-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002447_s2238-7854(12)70004-2-Figure4-1.png",
+        "caption": "Fig. 4 Test rig used in friction experiments 1: base; 2: inclined plane; 3: sample; 4, 5, 6: two disks and string, respectively; 7: protractor, indicator of the angle[8]",
+        "texts": [
+            " The rules of friction are not the same for instance in the press shown in Fig. 2 and in the silicon micromotor depicted in Fig. 3. This statement will be supported by the results obtained performing a very simple tribological experiment. The size of the rubbing/contacting surfaces (together with the magnitude of the applied load) has been decreased considerably from one test to another and the effects of these changes in the friction coef cient were observed. The test rig used to carry out the friction experiments[8] is presented in Fig. 4. The simple inclined plane to measure the static friction coef cient was very useful since the friction coef cient was estimated by the measurement of the angle of inclination of the plane to the horizontal plane. The gravity force was used to load the rubbing element. The aluminium samples (15 m thick foil, folded due to its large area) of the selected weight with rather low surface roughness were placed on the inclined polished Edi\u00e7\u00e3o 01.indb Art14 22/06/2012 18:09:58 steel plate (Fig. 4). Prior to the test both the sample surface and the inclined plane were carefully cleaned using cleaning solvents and nally by the use of petroleum spirit. The experiments started with the sample of gravity force 20 mN and additional weights up the total load 1.28 N. The area of contact of the foil was 4.76 x 104 mm2. After each test (a few slides have been realized with one weight of the sample) the load was decreased by half taking out the weights and nally the foil was cut off. The experiments nished within the area of the foil of 6 mm2 and at the load of 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003749_s1068366612020110-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003749_s1068366612020110-Figure6-1.png",
+        "caption": "Fig. 6. Schematic of calibration of bending load control system of SI 03M test machine.",
+        "texts": [
+            " Before the check, the contact loading system is calibrated. The load is assigned in the manual control mode. The fine adjustment of the load is car ried out using the rod (see Fig. 5). The measurement error of the contact load FN is \u22482%. (6) The error of the system of measuring the bend ing load is determined using the DOSM 3 1U master proving ring. The check involves the following two modes: \u2014when the bending load is directed \u201cUPWARDS\u201d; \u2014when the bending load is directed \u201cDOWNWARDS\u201d. The arrangement of the devices and proving rings is shown in Fig. 6. Before calibration, the bending load control system is calibrated. The load is assigned in the manual control mode. The fine adjustment of the load is carried out using the rod (see Fig. 6). The accuracy of measurement of the bending load is 2%. (7) The error of the friction torque measurement system is determined using the reference weights of the 4th class (GOST 1328\u201382). The arrangement of the device is shown in Fig. 7. The lever is initially equili brated with a weight and then the friction torque meter is halted using the rod. The check is carried out by placing reference weights of the 4th class with masses of 0.1, 0.2, 0.4, 0.6, and 0.8 kg\u2014corresponding to friction torques of 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000525_s00170-006-0874-y-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000525_s00170-006-0874-y-Figure1-1.png",
+        "caption": "Fig. 1 Simulation of tooth cylindrical gear generation with hobbing",
+        "texts": [
+            " Model surface of tooth flanks allows analyzing monomial flank deviations, as well as radial composite and runout ones. Modelling tooth flanks in the mechanical desktop environment makes use of commands of turning, copying and taking away the drawn solids of the wheel and the tools. The model of the tools was drawn across sweep of the tooth cutting flanks along a helical path and subtracted from a cylinder. The model of the wheel was a cylinder. The parameters of models corresponding to the parameters of the real tool and wheel. The kinematics of the hobbing simulation was presented in Fig. 1. Reading precision of coordinates of model surface was 0.1 \u03bcm. The reciprocal turning of the toothed wheel and the tool is not smooth but rather stroke-like in character. The interdependence be- J. Michalski (*) : L. Skoczylas Faculty of Mechanical Engineering and Aeronautics, Rzesz\u00f3w University of Technology, W. Pola 2, 35-959 Rzesz\u00f3w, Poland e-mail: jmichals@prz.rzeszow.pl tween a wheel revolution and the tool revolution is described with equation (1). i \u00bc \u03d5 2 \u03d5 1 \u00f01\u00de where i is a transmission ratio; \u03d5 2 is revolution of the tool corresponding to the wheel revolution by angle \u03d5 1 (\u00b0). The simulation procedure in a mechanical desktop environment, was written as follows: 1. Rotate gear wheel{revolution of generated gear about value \u03d5 1 (Fig. 1)} 2. Rotate tool{tool revolution corresponding to a wheel revolution about value \u03d5 2 (Fig. 1)} 3. Copy and move tool{tool copy and its placement on the turning diameter of the wheel, value-a, c (Fig. 1)} 4. Subtract gear wheel-tool{subtraction of the dipped tool solid volume from the gear wheel model} 5. Repeat{restarting the procedure}. The kinematics of the hobbing simulation was presented in Fig. 1. In the case of hobbing, the tool is a model of a hob. The computer simulation of the gear generation takes place in a three dimensional mechanical desktop environment. A tooth profile is formed due to the turning of the gear wheel and the tool. The helix results from the hob blade moving along the wheel axis. The rotary motion of the tool \u03d5 2 is dependent on the rotary motion of the generated gear wheel \u03d5 1 in accordance with the transmission ratio of the technological gear, equation (1). Shift c corresponds to the value of the axial feed per wheel turn"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.11-1.png",
+        "caption": "Figure 1.11 depicts the resulting study-state cornering motion. The vehicle side slip angle ,8 has been indicated. It is of interest to note that at low speed this angle is negative for right-hand turns. Beyond a certain value of speed the tyre slip angles have become sufficiently large and the vehicle slip angle changes into positive values. In Exercise 1.2 the slip angle fl will be used.",
+        "texts": [
+            "46) After elimination of the lateral velocity v we obtain the second-order differential equation for the yaw rate r\" Imui\"+ { I (C 1 +C2) +m(aZC1 +b2C2)} t~+ 1 +--{C1C212-muZ(aC1-bC2) } r - m u a C 1 3 + C 1 C z l 6 (1.47) U Here, as before, the dots refer to differentiation with respect to time, ~ is the steer angle of the front wheel and I (=a + b) represents the wheel base. The equations may be simplified by introducing the following quantities\" C - C l + C 2 Cs - Cla - Czb (1.48) Cq 2 - C1 a2+C2 b2 m k 2= I Here, C denotes the total cornering stiffness of the vehicle, s is the distance from the centre of gravity to the so-called neutral steer point S (Fig. 1.11), q is a length corresponding to an average moment arm and k is the radius of gyration. Equations (1.46) and (1.47) now reduce to: m( f +ur) + C Cs - - v + ~ r - C16 u u (1.49) cq 2 Cs mkZi '+ r+ v - Cla(~ U U and with v eliminated: mZkZuZkz+mC(q 2 +k2)u i .+(C1CzlZ-mu2Cs)r -mu2aCld+UC1Czlc~ (1.50) The neutral steer point S is defined as the point on the longitudinal axis of the vehicle where an external side force can be applied without changing the vehicle's yaw angle. If the force acts in front of the neutral steer point, the vehicle is expected to yaw in the direction of the force; if behind, then against the force",
+            " The side slip phase response tends to -270 degrees (at larger speeds) which is due to the negative coefficient ofjo9 in the numerator of (1.79). This in contrast to that coefficient of the yaw rate response (1.77). It is of interest to see that the steady-state slip angle response, indicated in (1.79), changes sign at a certain speexl V. At low speeds where the tyre slip angles are still very small, the vehicle slip angle obviously is negative for positive steer angle (considering positive directions as adopted in Fig. 1.11). At larger velocities the tyre slip angles increase and as a result, fl changes into the positive direction. Exercise 1.2. Four-wheel steer, condition that the vehicle slip angle vanishes Consider the vehicle model of Fig.l.16. Both the front and the rear wheels can be steered. The objective is to have a vehicle moving with a slip angle,8 remaining equal to zero. In practice, this may be done to improve handling qualities of the automobile (reduces to first-order system!) and to avoid excessive side slipping motions of the rear axle in lane change manoeuvres"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001954_13506501jet521-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001954_13506501jet521-Figure2-1.png",
+        "caption": "Fig. 2 The schematic representation of the experiment rig",
+        "texts": [
+            " 1 with the following parameters given in Table 1. The inner wall of the Proc. IMechE Vol. 223 Part J: J. Engineering Tribology JET521 \u00a9 IMechE 2009 at PENNSYLVANIA STATE UNIV on May 23, 2015pij.sagepub.comDownloaded from bearing was built from tin-based Babbitt alloy, which is soft, wear resistant and heat conducting, and the bulk material was 40Cr. The composition of the tin-based Babbitt alloy was Sn 85.13 per cent, Pb 0.024 per cent, Cu 5.84 per cent, and Sb 10.61 per cent. The test set-up is shown in Fig. 2. The shaft was supported by two preloaded angular contact ball bearings and the test bearing was centrally located between them. The test section of the rotor was ground to a fine surface finish, with an average surface roughness of 0.4 \u00b5m. The journal was driven with a 17.5 kW three-phase direct current motor with a no load operating speed of 3000 r/min and a thyristor direct current-motor speed controller with which infinite speed variation could be achieved. Lubricant was fed to the bearing from the top side of the bearing and recirculated by a one-way fixed-displacement pump"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002481_physreve.84.042901-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002481_physreve.84.042901-Figure1-1.png",
+        "caption": "FIG. 1. (Color online) Schematic diagram of our model system. An isotropic elastic rod is forced to rotate at frequency \u03c90 at its clamped base with the other end free in a viscous fluid of viscosity \u03b7.",
+        "texts": [
+            " It is mentioned that an open polymer can release its axial rotation of 4\u03c0 by rotating around its end once [7,8]. Geometrically, this refers to the well-known importance of the spinor representation of the rotation group [8,10,11]. However, the role of this writhing geometry in the driven dynamics of elastic filaments [12] appears to have been much less explored so far, which is the purpose of the present study. Specifically, we study an isotropic elastic rod that is axially rotated at one end at frequency \u03c90 with the other end free (see Fig. 1). This model system exhibits a rich variety of elastohydrodynamic phenomena and was first proposed and analyzed by Wolgemuth et al. [13]. They showed that a shape instability occurs at a critical frequency \u03c9c. For \u03c90 < \u03c9c, the rod remains straight and undergoes simple axial spinning (twirling), but for \u03c90 > \u03c9c the rod buckles and exhibits a combination of axial spinning and rigid-body rotation (whirling). Large-amplitude whirling in the stationary state was later demonstrated numerically [14,15]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002631_1.3625712-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002631_1.3625712-Figure1-1.png",
+        "caption": "Fig. 1",
+        "texts": [
+            " The effect of both external and internal damping on the transverse vibration of the shaft is investigated in detail. The phenomenon of gravitational resonance in horizontal shafts is also examined. Equations of Mot ion of a Rotating Shaft Consider a thin, initially bent, unbalanced shaft of unsymmetrical cross section mounted in rigid axially symmetric bearings and rotating with constant angular velocity S2 about the horizontal centerline, OA', of the bearings. A thin slice of the shaft in its displaced position is shown in Fig. 1. 0 represents the origin of orthogonal coordinate systems, OXYZ fixed in space, and OXUV rotating with the shaft with angular velocity (.l about OX. The directions OF, OU are parallel to the principal axes of the shaft cross section and initially coincident with OX, OF. E denotes the elastic (or geometric) center of the cross section, and C the mass center which in general will be different from E if the shaft is unbalanced. In the displaced position of the shaft at time t, the coordinates of E are y, z referred to the stationary axes OY, OZ, and u, v referred to the rotating axes OU, OV",
+            " Considering equation (4) first and seeking separable solutions of the form ti(x, t) = 4>(x) cos (u)l + a) the spatial function <j>(x) is seen to satisfy the equation d44> dx4 = k4 where mcok l = ~EL (6) (7) (8) ing in the planes OXU, O X V by Eh, Eh, respectively. The damping forces acting on unit length of the shaft are assumed to lie viscous and to consist of the following: 1 External (stationary) damping of magnitude ( \u2014 ma) times the transverse velocity of E measured relative to the stationary axes OXYZ. 2 Internal (rotary) damping of magnitude (\u2014m/3) times the transverse velocity of the cross section relative to the rotating axes OXUV. Under these assumptions the equations of motion of the thin slice, shown in Fig. 1, may be derived and expressed either in terms of the coordinates y, z referred to stationary axes OXYZ, or in terms of the coordinates u, v referred to rotating axes OXUV. In the present example, in which the shaft cross section is unsymmetrical, the equations in y, z will contain periodic coefficients while those in it, v will not. The latter pair of equations is therefore the easier one to handle and is found to be3 The general solution of equation (7) is 4>(x) = A cos kx + B sin kx + C cosh kx- + D sinh k"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002014_cefc.2010.5481434-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002014_cefc.2010.5481434-Figure3-1.png",
+        "caption": "Fig. 3. The prototype of M-FSPM machine.",
+        "texts": [
+            " Thus, the individual windings are essentially isolated among phases, leading to significantly enhance the FT capability. Fig. 2 shows the no-load magnetic field distributions. It depicts that the M-FSPM motor can offer the nature of phase decoupling, which is essential for FT operation. By using finite element method, the proposed M-FSPM motor has been quantitatively compared with the conventional FSPM motor. An experimental M-FSPM motor is designed and prototyped for verification, as shown in Fig. 3. This work was supported in part by grants (Project No. 60974060, 50907031 and 50807007) from the National Natural Science Foundation of China and a grant (Project No. 20080769007) from the Aeronautical Science Foundation of China. [1] S. Gopalakrishnan, A. M. Omekanda, and B. Lequesne, \u201cClassification and remediation of electrical faults in the switched reluctance drive,\u201d IEEE Trans. Ind. Appl., vol. 42, no. 2, pp. 479-486, Mar./Apr. 2006. [2] Z. Sun, J. Wang, G. Jewell, and D. Howe, \u201cEnhanced optimal torque control of fault-tolerant permanent magnet machines under flux weakening operations,\u201d IEEE Trans"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000036_ipemc.2006.4778074-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000036_ipemc.2006.4778074-Figure1-1.png",
+        "caption": "Figure 1. Magnetic pole identification based on saturation effects",
+        "texts": [
+            " The principle of the magnetic pole identification is briefly discussed and the proposed method introduced. The validity of the proposed identification method is verified by experiment results. II. IDENTIFING PERMANENT MAGNET POLARITY The principle of identifying the magnet pole based on magnetic saturation effects has been discussed in literature and is illustrated in Figs.1 (a), (b) and (c). Initially, the flux increases in direct proportion to the increase of currents. Further increases in currents result in progressively smaller increases in flux because of magnetic saturation. As shown in Fig.1 (c), with the permanent magnet excitation alone, the original operating point is \u201cA\u201d. However, with the stator excitation voltage applied in the same polarity with respect to that of the permanent magnet, the stator currents will increase from i0 to i1, and operating point from \u201cA\u201d to \u201cB\u201d due to the deep saturation. On the other hand, when the applied stator excitation voltage and magnetic pole are in the opposite directions, the magnetic path will not be saturated. The stator current will change from i0 to i2. At the same time, the operating point moves accordingly from point A to C. In these two cases, the applied volt-seconds are the same but different in polarity. Due to core saturation nonlinearity, the magnitudes of current variations are different. Therefore, we can detect the polarity of permanent magnets mounted on the rotor. When the magnetic pole is in an arbitrary initial position as shown in Fig.1 (a), we have to detect the axis of the permanent magnet first, normally achieved by the high frequency voltage injection method but the polarity of the magnet is left unknown. In the second step, according to the axis direction of the permanent magnet, positive and negative pilot voltages are applied sequentially in a controlled mode such that the voltseconds are the same in both cases. When the positive voltage is applied, the current incremental is +|i1| and when negative the current incremental is -|i2|. In the third step, we compare the magnitudes of the current incrementals. If |i1| is larger than |i2|, the magnetic pole that corresponds to |i1| is the North and the rotor position in the range from 0 to \u03c0 . Similarly, when the magnetic pole is in the orientation as shown in Fig.1 (b), the corresponding current (|i2|) with the positive stator excitation applied will be smaller than current (|i1|) with the negative stator excitation applied. Therefore, the magnetic pole that corresponds with |i1| is the North Pole. In this case, the rotor position angle is actually an angle between \u03c0 and \u03c02 and the angle will be \u03c0\u03b8 +r . III. IDENITFYING POLE POLASRITY BASED ON SPACE VECTOR PWM In the proposed magnetic poles identification scheme, the principles discussed above are applied but the voltages are based on the space vector PWM that is readily available in the controller hardware and software"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000073_4-431-31381-8_14-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000073_4-431-31381-8_14-Figure1-1.png",
+        "caption": "Fig. 1. Robot model",
+        "texts": [
+            " Moreover, the learning module is added to the state machine controller so that the minimum energy walking can be realized. The rest of the paper is organized as follows. First, the state machine controller to realize ballistic walking is introduced. Next, the learning module to optimize the parameters of the state machine controller is described. Then, the proposed controller is applied to a biped model that has the same length and mass to a human. Here, we use a robot model consisting of 7 links: a torso, two thighs, two shanks and two foots as shown in Fig.1. The parameters of the robot are shown in Table 1. The state machine controller at each leg consists of four states, as shown in Fig 2: the beginning of the swing phase (swing I ), the middle of the swing phase (swing II ), the end of the swing phase (swing III ), and the support phase (support). In the support phase, the hip joint is controlled with a proportional derivative (PD) manner so that the torso stands up and the support leg goes back. To the knee joint, torque is applied so that the knee joint becomes straighten during the support phase"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003932_1077546312461026-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003932_1077546312461026-Figure1-1.png",
+        "caption": "Figure 1. Ship coordinate systems.",
+        "texts": [
+            " The nonlinear six degree-of-freedom ship model, presented in Khaled and Chalhoub (2011) and briefly described herein, is used as a test bed to assess the performances of the proposed controller and observer. Its formulation incorporates recent advances in both maneuvering and seakeeping theories (Newman, 1977; Fossen, 1994, 2002, 2005; Fossen and Grovlen 1998; Journe\u0301e and Pinkster, 2002; Kristiansen et al., 2005; Perez, 2005). The model accounts for the surge, sway, heave, roll, pitch and yaw motions of the ship (see Figure 1). Following the Society of Naval Architects and Marine Engineers (SNAME) (1950) convention, both position and orientation of the ship are defined with respect to the inertial frame while the ship at NATIONAL CHUNG HSING UNIV on April 12, 2014jvc.sagepub.comDownloaded from translational and angular velocity vectors are determined with respect to the body-fixed frame. The equations of motion are given in Khaled and Chalhoub (2011). The waves are considered to be long-crested with a Modified Pierson-Moskowitz spectrum (Perez, 2005)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure24-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure24-1.png",
+        "caption": "Fig. 24. Rivet deformations a) w1 b) w2",
+        "texts": [
+            " Using this procedure, the iteration process is done simultaneously to satisfy both the contact constraints and global equilibrium using the Newton-Raphson method. This procedure is accurate and stable but may require additional iterations. Numerical analysis The numerical calculations were performed for two cases of upsetting (w1, w2) distinguished by the height of the formed rivet head (Table I, Fig. 21). Methods for FEM analysis of riveted joints of thin walled aircraft structures 959 Riveted joint deformations are shown in Fig. 23, while rivet deformations for two cases of upsetting are given in Fig. 24. In the first case (w1), the height of formed rivet head has maximum value, when the diameter reaches nominal value [17]. For the case w2, diameter of formed rivet head reaches maximum permissible value according to riveting process manual [2]. Rivet upsetting causes metal sheet joining and filling of the rivet hole. Because of increasing fatigue life of the riveted joint, the rivet upsetting and compressive stresses distribution in the sheets around the hole should be uniform and exceed yield stress level [18,19]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001952_ichr.2009.5379580-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001952_ichr.2009.5379580-Figure9-1.png",
+        "caption": "Fig. 9. A visually supported execution. The red lines represent the displacement of the original position on the path and the visually determined poses in workspace.",
+        "texts": [
+            " 61q = ComputeCSpaceDelta(e,q) (6) Algorithm 1: C omputeCSpaceDelta(61x, q) 1 Xs tart f- ForwardKinematics(q) ; 2 q' f- q;61x' f- 61x ; 3 while (fTimeOut()) do 4 61q f- kJ+(q')61x'; 5 q' f- q' + 61q; 6 Xn ew f- ForwardKinematics(q'); 7 61x' = Xn ew - Xs tart ; 8 if (1161x' - 61xll :::; threshold) then 9 return (q' - q); 10 end By moving the C-Space path by 61q during trajectory execution, the real TCP pose is adjusted so that the error in the relation between target and hand is reduced. The resulting trajectory will not be valid in the virtual planning surround ing any more (see Fig. 9), but in reality the displacements of the TCP can be reduced and a more precise execution is possible. Since 61q is calculated in each loop, the influence of static and dynamic errors is decreased. F. Including Reactive Components Additionally to the proposed algorithms, the robot must be able to register unplanned contacts with the environment or humans and the system must react accordingly. Also interactions by a human operator, i.e. a correction of an executed trajectory, must be recognized and processed",
+            " The task includes the recognition and localization of the target object and obstacles to build up the internal representation of the scene (see sec. III), which then is used for planning a collision-free motion to reach the object with either the left or the right hand. Searching two paths in parallel is done by the MultiEEF-RRT planner as described in section IV. If a solution is found by the planner, the proposed reactive execution framework is parametrized and the planned trajectory is reliably executed. In Fig. 9 a planned trajectory is shown in workspace with white dots and the black line. During execution the visually determined hand position is used to calculate the displacement e, which is visualized by red lines. As shown the error between virtual world and reality changes during the execution of the trajectory. The workspace positions of the temporary goals are depicted as red dots and due to the Jacobian based projection of the error from workspace to C-Space the displacement is not exactly represented, but a good approximation is served by the algorithms. The recognized TCP position with respect to the target object is marked by the green path which is close to the ideal planned trajectory. The large deviations between the virtually planned and the executed trajectory, depicted as red lines in Fig. 9, are mainly caused by an inaccurate calibration of the joint sensor data. The joint position is derived from incremental sensors at the driving end, resulting in approximated joint values used as input for the control loop. The accumulated error of seven arm joints, one hip joint and three joints of the robot's neck causes large workspace displacements. Furthermore the active head is moving due to the actuated hip and neck joints. This means that a moving camera system produces a varying error in the transformation between the visually determined object poses and the global coordinate system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001831_kem.442.130-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001831_kem.442.130-Figure1-1.png",
+        "caption": "Fig. 1 (a) GCr15 ring and (b) the surface area to be hardened.",
+        "texts": [
+            " All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-02/06/15,01:38:40) The material was received in the form of rings with hardness of 300-350Hv (30-33 HRC). The material already had conventional heat treatment to achieve this hardness. The drawing of the one such ring from a pair of rings is shown in Fig. 1. The main objective was to harden the surface A of the ring (as shown in Fig. 1) up to a required hardness of 750-850 Hv (60-65 HRC). A transverse flow CO2 laser was used to heat the required surface of the ring. The experimental set up for laser treatment is shown in Fig. 2. A rectangular laser beam of required size of 3.5 x 2.5mm2 was used to heat the work piece. The focusing mirror used had a focal length of 280 mm. In order to increase the absorption of the laser, the surface of the ring to be hardened was coated black. For laser hardening, the samples were mounted on a jig in CNC chuck which rotated under the laser beam"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure13-1.png",
+        "caption": "Fig. 13 An inertially symmetric rigid-body undergoing spherical motion",
+        "texts": [
+            " APPENDIX 1 Notation a vector of the dimension indicated in the text A m6n matrix, with m>n, m, and n as indicated in the text A{ left Moore\u2013Penrose generalized in- verse of the above matrix A: A{5 (ATA)21AT (A{) time derivative of A{ CPM(v) cross-product matrix of vector v, usually indicated as V J limb number, a Roman numeral, for J5 I, II Omn m6n zero matrix, with m and n indicated in the text Tb twist-shaping matrix of body b map- ping actuated joint rates into bodytwist T\u0304b twist-shaping matrix of body b map- ping motor angular velocities into body-twist 0n n-dimensional zero vector, with n indicated in the text 1nn n6n identity matrix, with n indicated in the text hJi ith angle of rotation of limb J, measured as indicated in the text h four-dimensional vector of actuated joint variables q four-dimensional vector of motor angular displacements vi scalar angular velocity of the ith element in the epicyclic gear train (EGT); also a component of the angular velocity vector moving in three-dimensional space ||?|| Euclidean norm of vector (?) Proc. IMechE Vol. 223 Part I: J. Systems and Control Engineering JSCE623 F IMechE 2009 at Purdue University on June 28, 2015pii.sagepub.comDownloaded from APPENDIX 2 The NOC modelling of a rigid-body undergoing offset spin and precession Theorem 2 is proved here. To this end, the reader is referred to Fig. 13, showing an axially symmetric body, namely a frustum, to suggest the shape of the pitch surface of bevel gears. However, the shape of the body under study can be more general than this, as long as it is axially symmetric. It goes without saying that the body is assumed to be made of a homogeneous material, which is uniformly distributed throughout the body. In Fig. 13, the body is shown to spin about its axis of symmetry S and to precede about a vertical axis V, which is fixed to an inertial frame, the two axes being skew. Let e, f, k be a triplet of orthonormal vectors in which k5 e6f. Notice that the axis of symmetry and the axis of precession do not necessarily intersect, in order to keep the layout as general as possible. Also note that, in the case of bevel gears 9 of the RAGB, d5a, vk5vp, and ve5vt. Thus, t takes the form t~ veezvkk veezvkk\u00f0 \u00de| aezbf\u00f0 \u00de \" # ~ k e af{be bk \" # |fflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflffl} T [ R6|2 vk ve \" # \u00f091\u00de thereby obtaining T"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001599_19346182.2009.9648507-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001599_19346182.2009.9648507-Figure3-1.png",
+        "caption": "Figure 3. Coordinate systems; sensor: top row; ball: bottom row.",
+        "texts": [
+            " During the assembling of the ball, additional weight blocks were added to restore the original weight of the ball and to accommodate for heavy ball strokers and crankers, and light ball spinners. It is important to notice that the transducers are connected to interface power supply boxes by cables. These cables did not compromise the experiments, as, once the ball is released, the fingers no longer applied forces to the ball and thus the ball was not required to roll or skid after release and could be stopped by foam cushions. The orientation of the sensor coordinate systems is shown in Figure 3. The coordinate system of the ball is: x-axis Figure 1. Design of the instrumented bowling ball (Pro/ENGINEER, different views); 1: connectors; 2: 6D-transducers; 3: thumb tube; 4: finger tubes; 5: base plate; and 6: screws (connecting bottom and top half of the ball). www.sportstechjournal.com & 2009 John Wiley and Sons Asia Pte Ltd Sports Technol. 2009, 2, No. 3\u20134, 97\u201311098 D ow nl oa de d by [ C hu la lo ng ko rn U ni ve rs ity ] at 2 2: 48 1 1 Ja nu ar y 20 15 pointing from ring and middle finger towards the thumb, y-axis from the centre of the ball to a point between the openings of the finger and thumb holes, and z-axis from ring to middle finger",
+            " The latter filter, however, shortens the moment spikes and alters the initial and final segments of the data set. The forces (Fx, Fy, Fz) and moments (Mx, My, Mz) of the thumb, middle and ring finger were processed as follows: The resultant finger force Fd is calculated from Fd \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Fx2d \u00fe Fy2d \u00fe Fz2d q \u00f01\u00de where the subscript d denotes either the thumb, the middle or the ring finger. After rotating the sensor coordinate system about the x-axis by 1291, i.e. the inclination angle of the tubes (Figure 3), Fy0Th \u00bc FyTh cos \u00f029 \u00de \u00fe FzTh sin \u00f029 \u00de \u00f02\u00de Fz0Th \u00bc FyTh sin \u00f029 \u00de \u00fe FzTh cos \u00f029 \u00de \u00f03\u00de Fy0MiRi \u00bc \u00f0FyMi \u00fe FyRi\u00de cos \u00f029 \u00de \u00fe \u00f0FzMi \u00fe FzRi\u00de sin \u00f029 \u00de \u00f04\u00de Fz 0 MiRi \u00bc \u00f0FyMi \u00fe FyRi\u00de sin \u00f029 \u00de \u00fe \u00f0FzMi \u00fe FzRi\u00de cos \u00f029 \u00de \u00f05\u00de Figure 2. Final prototype (same legend as Figure 1). Sports Technol. 2009, 2, No. 3\u20134, 97\u2013110 & 2009 John Wiley and Sons Asia Pte Ltd www.sportstechjournal.com 99 Instrumented bowling ball D ow nl oa de d by [ C hu la lo ng ko rn U ni ve rs ity ] at 2 2: 48 1 1 Ja nu ar y 20 15 where the subscripts Th, Mi, and Ri denote thumb, middle and ring finger respectively",
+            " The spin shot, although the terminal negative z-spike is still present, is dominated by the positive Myball peak in the second half of the forward swing. The instrumentation of sports equipment is often constrained by the rules issued by the governing sporting bodies [11]. According to the bowling rules, a tenpin bowling \u2018ball shall be constructed of solid material ywithout voids in its interior and be of a non-metallic composition\u2019 [12]. Consequently, an Figure 10. 3D vector diagrams (straight shot, beginner); the coordinate system indicated corresponds to the ball coordinate system (c.f. Figure 3). The Force Vectors of back and forward swing are represented as single vectors; the scale bar corresponds to 20 cm (ball dimensions) and 40 N (force vectors); the force vectors applied by the ball to the hand are displayed on the ball and the finger tubes (where the forces originate from); the first three rows show the entire vector diagram; the 4th row shows selected force vectors; Th 5 thumb, Mi 5 middle finger, Ri 5 ring finger. The Moment Vectors of the forward swing are represented as a regular surface; the scale bar corresponds to 1 m (ball dimensions) and 1 Nm (moment vectors) in the right view, and to 5 m and 5 Nm in the top and front views; the origin of the moment vectors is set to the centre of the ball"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001417_s00791-008-0109-x-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001417_s00791-008-0109-x-Figure5-1.png",
+        "caption": "Fig. 5 Families of semipermeable curves for a = 0",
+        "texts": [
+            " Every curve of the family (2),2 terminates on the interval (\u2212\u221e,\u2212a] of the horizontal axis. Those curves of the family (2),1 which come to the interval (\u2212a, a) of the horizontal axis are smoothly continued by the curves of the family (2),2. The other curves of the family (2),1 terminate on the interval (\u2212\u221e,\u2212a] of the horizontal axis. If a = 0, families (1),1, (1),2 of semipermeable curves of the first type and families (2),2, (2),1 of semipermeable curves of the second type can be introduced as well (see Fig. 5). If a < 0, the velocity vectogram of system (11) contains zero in its interior for any point z with y = 0. Therefore, vectors f (1)(z), f (2)(z) do not exist. Hence, for a < 0, semipermeable curves are absent. The set W(\u03c4, M) is a solvability set in the problem of reaching a given set M by time \u03c4 for system (11). For any \u03c4 \u2217 > \u03c4\u2217, we have W(\u03c4 \u2217, M) \u2283 W(\u03c4\u2217, M). Take any z in the plane of phase variables of system (11). Put V (z) := min{\u03c4 : z \u2208 W(\u03c4, M)} if there exists such \u03c4 \u2265 0 that z \u2208 W(\u03c4, M)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003005_s0263574711001172-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003005_s0263574711001172-Figure1-1.png",
+        "caption": "Fig. 1. The 3-RPR planar parallel manipulator.",
+        "texts": [
+            " The roles of learning, evaluating, testing, and fine tuning the SOM are demonstrated in Section 4, in which the concept of tackling the direct determination of the unique solution among the assembly modes is presented. Simulations are conducted in Section 5 to show the performance of the developed method and bear out its efficacy. A comparative study is also carried out with the numerical NR technique. Conclusions are finally drawn in Section 6. The simulated manipulator considered here is the 3-RPR planar parallel manipulator with moving platform defined by moving pivots Mf and a fixed platform defined by fixed pivots Ff , f = 1, . . . , 3, as shown in Fig. 1. Point G is a point in the moving platform. The vectors of the active prismatic joints of the three limbs connecting the base and the moving platforms are lf for f = 1, . . . , 3. The orientation angles of l f with respect to a fixed coordinate frame (X, Y) at F1 are \u03b8f ; those angles are the passive joint angle variables. For the forward kinematic problem of this manipulator, the given vector is the active prismatic joint variables vector defined by g = [ l1 l2 l3 ]T , while the required vector is the end-effector position and orientation (pose) vector defined as x = [xG yG \u2205 ]T . The passive joint angle variables vector is given by \u03b8 = [ \u03b81 \u03b82 \u03b83 ]T . The kinematic analysis of such manipulator reveals that the inverse kinematics has unique solution while the forward kinematics has at most six possible solutions.20\u201322 The differential kinematics is given below. The loop closure equation for the first limb, as depicted in Fig. 1, is given by: r1 + e1 = l1, (1) where r1 is the position vector of point G and e1 is a vector pointing from point G to the moving pivot M1. The time derivative of Eq. (1) yields the velocity vector-loop equation of the first limb as follows: vG + \u2205\u0307(k \u00d7 e1) = l\u03071u + l1\u03b8\u03071(k \u00d7 u), (2) where u = cos \u03b81 i + sin \u03b81 j + 0k is a unit vector along F1M1, in which i , j , and k are unit vectors in the positive X-, Y-, and Z-directions, respectively. vG is the velocity of point G. In order to get a scalar equation, both sides of Eq",
+            " Training the SOM network takes place through the procedure described in Section 3.2. The learning rate \u03b7(t) starts with \u03b70 = 0.2 and gradually decreases up to 0.01 with the time constant \u03c42 = 1000. The effective width parameter of the Gaussian function \u03c3 (t) starts with \u03c30, which is set equal to the radius of the lattice, and the time constant \u03c41 = 1000/log\u03c30. This manipulator is symmetric with three identical limbs connecting the equilateral triangle of the base and that of the moving end-effector as shown in Fig. 1. Each side of the moving equilateral triangle is 100 mm, while that of the base is 300 mm. The minimum and the maximum lengths of limb f are lf min = 100 mm and lf max = 300 mm for f = 1, . . . , 3, respectively. Point G is taken as the median point of the moving end-effector. The given end-effector trajectory is specified as a straight line trajectory in the two-dimensional space starting at point \u201cA\u201d (120, \u221213.4) mm with orientation angle \u2205 = 60\u25e6 and ending at point \u201cE\u201d (120, 136.6) mm with the same orientation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003710_gt2012-68956-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003710_gt2012-68956-Figure5-1.png",
+        "caption": "FIGURE 5. Tested brush core elements",
+        "texts": [
+            " Therefore, a sensing device, which is connected to a load cell, can be moved into the bristle pack while monitoring both force and deflection. Contrary to Bidkar et al. [6], the test specimen of the Braunschweig test rig is a finite-extent shoe. It is well known that this testing setup leads to results containing endeffects. However, since the test rig is used for comparing brush seals prior to and after tests in steam environment, the absolute stiffness is of minor importance. The tests described here were conducted with two different types of brush seals (see FIGURE 5). First, a standard brush seal was used. This design is characterized by a bristle pack which is leaned against the backing plate, i.e. the bristle pack is inclined in axial direction. 3 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/74826/ on 01/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use cross section of the brush core element FF gapping slotted tube core wire bristles gapping back plate front plate back plate front plate bristlespack FIGURE 6"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000420_00022660810859346-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000420_00022660810859346-Figure1-1.png",
+        "caption": "Figure 1 The 2DOF helicopter apparatus",
+        "texts": [
+            " From equation (25), if the following condition holds: _g ii lmin \u00f0Qi\u00de . Hiilmax \u00f0Pi\u00de \u00f026\u00de _Vi , 21i \u00f027\u00de where 1i is some finite constant which implies the stability is guaranteed. A In this section, a two degrees of freedom helicopter CE150 from Humusoft (1989) was used in this experiment. The model is a multidimensional, nonlinear system with two manipulated input and two measured outputs (horizontal and vertical angles) with significant cross couplings. The system consisted of a massive support, and the main body, carrying two propellers driven by DC motors, Figure 1. The model can Direct adaptive fuzzy control for a nonlinear helicopter system H.F. Ho, Y.K. Wong and A.B. Rad Volume 80 \u00b7 Number 2 \u00b7 2008 \u00b7 124\u2013128 be described by the nonlinear state equations with six states, two inputs, which are the control value for main and side propeller motors. The two output values are the elevation and azimuth angles. The dynamics of the helicopter are represented by the following stated space model: _x1 \u00bc x2 _x2 \u00bc 1 Im \u00f0K1x3jx3j2 \u00f0b11jx3j \u00fe b21\u00dex2 2 Tc1 sign\u00f0x2\u00de\u00f012 e2\u00f0jx2 j= _u01\u00de\u00de 2 Tgsin\u00f0x1 \u00fe a\u00de \u00feKGx5x3cos\u00f0x1\u00de \u00feKcx 2 5 sin\u00f02\u00f0x1 \u00fe a\u00de\u00de\u00de _x3 \u00bc 1 Ir1 \u00f0u1 2 a1x3jx3j2 b1x3\u00de _x4 \u00bc x5 _x5 \u00bc 1 Is \u00f0K2x6jx6j2 1 sin\u00f0x1\u00de \u00f0\u00f0b12jx6j \u00fe b22\u00dex2 2 Tc2 sign\u00f0x5\u00de\u00f012 e2\u00f0jx5 j= _u02\u00de\u00de \u00fe sin\u00f0x1 \u00fe c\u00de sin\u00f0x1\u00de Kr1\u00f0u1 2Kr1\u00f0a1x3jx3j \u00fe b1x3\u00de\u00de\u00de _x6 \u00bc 1 Ir2 \u00f0u2 2 a2x6jx6j2 b2x6\u00de \u00f028\u00de All the constant numbers are given in the Humusoft (1989)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001447_robot.2007.363852-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001447_robot.2007.363852-Figure2-1.png",
+        "caption": "Fig. 2. Friction during knotting / unknotting",
+        "texts": [
+            " We will study the static friction and focus on how to tie a knot tightly and unknot a tight knot in the future. During the simulation, we use Coulomb's model and consider each suture segment as rigid body, hence we can not bend to any angle for any instant time. From Coulomb's observations we know that: kinetic frictional force is approximately independent of contact area and velocity magnitude of the object; Coefficient of friction depends on pairs of materials. During knotting or unknotting procedure, suppose there are only two segments colliding with each other (see Fig. 2). Let g be the friction constant, e be the friction direction vector, n be the force of repulsion, then the friction ff can be described as: ff =gu nl e. (1) To calculate the repulsion force n, we introduce a springdamper between the contact point C and the end point E. n = (k,s(2r -d) -krd(vr n))n. (2) where kr, is a spring constant for the repulsion force, r is the radius of the suture model, d is the distance between point C to point E (see (b) of Fig. 2), krd is the damper constant for the repulsion force, Vr is the relative velocity of point C with respect to point E, -n is the unit vector from point E to point C. We use linear interpolation to compute the velocity of a point on the segment. For example (see Fig. 2), vc =(1 -a)va + avb, where a is the fraction of point C along PaPb, ve = (1 -b)vc + bVd, where b is the fraction of point E along PcPd. Then the relative velocity vr = Vc-Ve, and the friction direction vector e is computed as follows: a = X-arcsin( 1ei-i x ei ). The torsional spring force can be computed as follows: (Vrfk )nf- Vr e =vrfn Vr ||~(Vr n)n-Vr (3) B. Internal Forces 1) Linear spring force: The linear spring force is computed by comparing the current segment length, li, between point, Pi and Pi+l, with the rest length of the segment ir, and by projecting the resulting difference on the direction from point Pi to Pi+ ",
+            " When two suture segments are detected to be at a distance d < 2r from each other, then, an equal (but opposite) displacement vector is applied to each segment along. This displacement is just long enough to take the segments out of collision, with a slight \"safety margin\". Hence, each node is shifted away by r-d 2 + \u00a3/2. If a collision occurred, during real time simulation, we need to compute new velocities of mass points which are in'volved in the collision. Similarly to the method presented in [17], we apply impulses to the end points of these two segments. See Fig. 2 for the case where point C with relative position a along the segment PaPb interacts with point E with relative position b along the segment P.Pd Let i be the impulse, then, i = nAt. where n is the repulsion force that we can obtain from equation (3). Then we can compute the new velocities as follows: Let fp be the component of the input force fh along the segment direction, and fp is the input force propagated to point Pi from point P7+j. fp and f10 can be obtain from the following equations: fp = (f7 ei)ei fni (fhuem)eCn"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003576_016918611x603837-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003576_016918611x603837-Figure1-1.png",
+        "caption": "Figure 1. Multifingered hands and objects: (a) 2-D case, (b) 3-D case and (c) hand grasping an box.",
+        "texts": [
+            " It is adaptive to various polygonal objects to be manipulated, to various contact states between the fingers and the object, and to any number of fingers in the hand. As a result, the proposed numerical method provides a good and complete solution to the problem of the workspace of multifingered manipulation or coordination of multiple robots. We also point out that the optimization-based numerical framework or methodology for workspace generation may be widely applied to parallel robots and other sophisticated robotic systems, including humanoid robots [23]. Consider the grasp of an object using a multifingered hand (Fig. 1). Assume that the hand has m fingers, with n joints in total. Denote the joint angles by a vector D ow nl oa de d by [ N ov a So ut he as te rn U ni ve rs ity ] at 1 2: 52 1 0 Ja nu ar y 20 15 2296 Y. Guan et al. / Advanced Robotics 25 (2011) 2293\u20132317 = (\u03b81, \u03b82, . . . , \u03b8n). is the parameter describing the hand configuration. The hand structurally consists of tips and links of the fingers. Topologically, we regard a fingertip as a point t and a link as a line segment l, and call them the topological features of the hand. Then, the set of topological features for the hand is defined as H = {l1, l2, . . . , ln, t1, t2, . . . , tm}, where li (i = 1,2, . . . , n) and tj (j = 1,2, . . . ,m) represent the ith link and the j th fingertip, respectively. In the 2-D or planar case (Fig. 1a), the configuration of an object may be described by the position (x, y) and orientational angle \u03c6 of the object frame O, with respect to the palm frame P . Topologically, a polygonal object consists of features including vertices and edges. Denote the sets of vertices and edges of the object by V = {vi} and E = {ej }, respectively, where vi and ej represent the ith vertex and j th edge in turn. Each edge has two vertices. We use ei(vj , vk) to denote the edge formed by two vertices vj and vk . The set of topological features for the object is therefore O = {V,E}. In the 3-D or spatial case (Fig. 1b), the configuration of the object is described by the position (x, y, z) and orientation (\u03b1,\u03b2, \u03b3 ), in RPY or Euler angles. A polyhedral object consists topologically of faces in addition to vertices and edges. Similar to the sets of edges and vertices (E and V ), the set of faces of the object is denoted by U = {uk}, where uk represents the kth face. (We do not use F and fk here for face feature, since we reserve them for force notation later on.) A face contains several vertices and edges. We use ui(v1, v2, ",
+            " In the example, we generate the workspace of a hand grasping and manipulating a rectangular object of size 70 \u00d7 50 mm2. Provided that the hand\u2013object system is in a vertical plane, the object weight must be balanced. The hand is assumed to consist of two fingers with 2-d.o.f. each. Suppose the lengths of links are l1 = 50, l2 = 40, l3 = 50, l4 = 40 and d = 60 (d is the distance between the two finger base points) (all in mm). Let \u03b81, \u03b83 \u2208 [\u03c0/6,2\u03c0/3] and \u03b82, \u03b84 \u2208 [0,5\u03c0/12] (in rad); their positive directions are shown in Fig. 1a. Assume that the limits of driving torques of the hand are T1 = 400, T2 = 500, T3 = 500 and T4 = 400 (all in Nmm); the normal contact forces are f n 1 10 N, f n 2 10 N; the coefficient of friction is \u03bc = 0.3. In the manipulation, the two fingertips t1 and t2 contact with the two middle points, q1 and q2, of the left and right short edges, respectively, no sliding is allowed. Then the fixed-point contact constraints can be expressed as (xt1 \u2212 xq1) 2 + (yt1 \u2212 yq1) 2 + (xt2 \u2212 xq2) 2 + (yt2 \u2212 yq2) 2 = 0, where the coordinates can be easily obtained by finger kinematics and object configuration",
+            " It was found that the workspace in this case is much larger than those with fixed-point contacts [28]. The result is consis- D ow nl oa de d by [ N ov a So ut he as te rn U ni ve rs ity ] at 1 2: 52 1 0 Ja nu ar y 20 15 Y. Guan et al. / Advanced Robotics 25 (2011) 2293\u20132317 2311 tent with the prediction in Ref. [17] that if the contact points are allowed to move across the surfaces through rolling or sliding, it is possible to increase the size of the workspace greatly. The example is modeled according to a three-fingered hand manipulating a box of 35 \u00d7 25 \u00d7 75 mm3, as shown in Fig. 1c. The hand consists of three identical fingers with 3 d.o.f. each (finger f1 is hidden by f2 in Fig. 1c). The axes of the two outer joints in one finger are parallel and perpendicular to that of the first joint. The link lengths of one finger are 18, 50 and 40 mm, respectively. The finger coordinate frames are located at (7.2,30,19.2), (7.2,\u201330,19.2) and (7.2,0,\u201319.2), respectively, in the palm frame O\u2013XYZ, with angles of 22.5\u25e6 between their Xiaxes and the palm X-axis. The joint ranges of the first joints of the fingers are: \u03b81 \u2208 [\u221245,60], \u03b84 \u2208 [\u221260,45] and \u03b87 \u2208 [\u221260,60], and those of the second and third joints are \u03b82, \u03b85, \u03b88 \u2208 [\u221260,90] and \u03b83, \u03b86, \u03b89 \u2208 [0,90] (all in deg)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003515_j.procs.2012.06.155-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003515_j.procs.2012.06.155-Figure1-1.png",
+        "caption": "Figure 1: (a)Illustrative example with non-holonomic constraint. (b)Optimization problem to orientation sensing",
+        "texts": [
+            " The control signal saturation is implemented defining minimum and maximum values, Ui = cmin and U\u0304i = cmax. The ACvO (Algorithm of Consensus Via Optimization) only calculate the consensus trajectory associated with the information state, which is the understanding of each vehicle about the meeting point. We assume the use of local controllers to ensure that the vehicle reaches the desired position at every sampling time. It does not consider possible mechanical constraints in the motion of the vehicle, e.g., the orientation constraints of non-holonomic mobile robots. Figure 1(a) shows an illustrative example, where a vehicle with local controller, even following the trajectory can fail in meeting non-holonomic (lower trajectory). The vehicle orientation, at time k, is opposed to the direction of the next point, and thus, because of non-holonomic constraint, the vehicle makes a turn (upper trajectory), delaying its route to the consensus point, needing one more iteration to reach the trajectory. Moreover, assuming that the vehicle has a fixed camera, in which the sensing range can be associated to the vehicle orientation. We added an optimization routine on the sensing range motivated by the knowledge of the future control sequence (optimization of Ji) and hence, all points of the trajectory. The optimization of vehicle orientation can be performed by minimizing the squared error, Figure 1(b). Let the cost function: J\u03b8i = Np\u22121\u2211 k=1 \u2016(\u03b8i[k] \u2212 \u03b8i[k + 1])\u20162 , i = 1, ..., n (12) The goal is to minimize J\u03b8i , where some constraints can be imposed, such as the maximum individual rotation \u03b4d and curvature radius of the vehicle rc. Thus, the new information states associated to coordinates x and y are: \u03bex i [k + 1] = \u03bex i [k]cos(\u03b8i + \u03d5) \u03be y i [k + 1] = \u03bey i [k]sin(\u03b8i + \u03d5) (13) where \u03d5 is the rotation related to global reference, since \u03b4d and \u03b8i are local variables, Figure 1(b). Note that each iteration of the ACvO, the maximum rotation is 3\u03b4d (due to mechanical constraints and saturation of the control signal) and the final value of \u03b8i is defined by optimizing J\u03b8i . Based on the blocks diagram, Figure 2(a), at each ACvO iteration, the optimization of Ji generates the trajectory to Np horizon based on the information exchanged, and more, with the knowledge of all trajectory prediction, the optimal orientation of vehicle ith is also calculated (J\u03b8i ). However, only the first point is implemented and we assume that the vehicle has a controller and local sensing to achieve this local target point"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003560_j.talanta.2013.05.019-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003560_j.talanta.2013.05.019-Figure1-1.png",
+        "caption": "Fig. 1. Experimental set up used for the ET-RTP measurements.",
+        "texts": [
+            " After homogenization of the mixture, the polymerization process takes place and the mixture was left to dry for two weeks in the absence of light, until a constant weight is achieved. The dry xerogel was crunched and fragmented in an agate mortar and particle sizes of diameters between 80 and 160 \u03bcm were selected by sieving for further experiments. The two carriers employed in order to make the sample pass through the sensing phase were 3 M HCl and H2O, at a flow rate of 1 mL min\u22121. For the preparation of the calibration graphs, 1 mL of bromate standard (or sample) was injected into the flow system via an injection valve located in the H2O carrier channel (see Fig. 1). After injection, the standard is mixed with 3 M HCl by using a \u201cY\u201d type connector. Finally, the mixture passed through the measurement flow cell containing the sensing phase, giving rise to a transient signal as consequence of the energy transfer in presence of bromate ions. Reagent blanks were prepared and measured following the same procedure (in this case, it was introduced Milli-Q water instead of bromate standards). The proposed ET-RTP method was evaluated for the determination of bromate ions in different type of commercial flours by following the general procedure after a pre-treatment step"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003859_gt2013-95074-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003859_gt2013-95074-Figure3-1.png",
+        "caption": "Figure 3: Pivot geometry parameters",
+        "texts": [
+            " The key parameters of the analysis are listed in Table 1. The bearing is mounted in load between pivot position. The oil is supplied to spray-bars between the pads via an outer annular channel. Each spray-bar contains 19 nozzle bores with a diameter of 2.4 mm. The flow rates are further reduced by sealing baffles with a cylindrical bore of a diameter of 502 mm. A drawing of the bearing is shown in Figure 2. The identification of the pad temperatures is enabled by means of 100 thermocouples located 5 mm behind the sliding surface. With reference to Figure 3, the pads feature a radius of R2 = 60 m in axial direction on the back, which enables tilting in axial direction additionally in order to prevent larger influences of misalignment between the bearing and the shaft. Due to the small elliptical area of the contact between the pad and the liner, this pivot is very flexible. The inner radius of the liner is R3 = 322.5 mm and the pad back radius in circumferential direction is R1 = 287 mm. The back of the pads are directly in contact with the liner in the middle plane of the bearing, as the pads are radially fixed by holding pins on the leading edge and the spray bars on the trailing edge"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003905_s00542-013-2023-5-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003905_s00542-013-2023-5-Figure4-1.png",
+        "caption": "Fig. 4 Fluid thickness in the ridge or groove regions in terms of the circumferential coordinate",
+        "texts": [
+            " The curvature of a film in a journal bearing was neglected. Since film thickness is significantly less than bearing radius, the fluid film can be unwrapped into a plane. For a grooved bearing, the reynolds equation for the steady state, laminar, isothermal, and incompressible 1 3 flow is (Vijayaraghavan and Keith 1989; hashimoto et al. 2012). where the coordinate system (\u03c6, y) is fixed to the bearing. Fluid thickness in the ridge or groove regions in terms of circumferential coordinate is shown in Fig. 4 or The pressure field is continuous in the circumferential direction and the cavitation algorithm is based on reynolds condition When eq. 1 has been solved for the pressure in the equilibrium state, the radial and tangential loads can be obtained by integrating the pressure over the bearing area The load capacity can be derived as In dimensionless form, the load capacity is In analyzing the groove design, the pressure distribution within the fluid film was derived by solving the reynolds equation using a spectral element method"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000433_1.2736707-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000433_1.2736707-Figure2-1.png",
+        "caption": "Fig. 2 Cross-sectional structure of the microbearing tester",
+        "texts": [],
+        "surrounding_texts": [
+            "om: http://tribology.asmedigitalcollection.asme.org/ on 01/29/2016 Terms allows the microbearing tester to be operated below a specified amplitude of rotor oscillation. Note that the measurement position of the radial displacement is located outside of the two journal bearings, hence the measured amplitude of the radial rotor oscillation can be larger than the bearing clearance when the rotor is oscillating in the conical mode. The rotational speed is measured using an optical fiber displacement sensor. The rotor is made of Ti\u20136Al\u20134V, and the surface of the shaft is coated with 2- m-thick CrN. The rotor weighs 2.3 g, and is dynamically balanced to the level that the residual unbalance is less than 0.1 g m, which corresponds approximately to JIS B 0905 G6.3 grade. The fabricated rotor is shown in Fig. 4. The rotor is designed to have sufficient margin between the rated rotation speed and the resonance speed of the first bending mode, which appears at 119,000 rpm. The bearing sleeves are a half-splitting type as shown in Fig. 5. By adopting the half-splitting type, the microbearing tester can be assembled without reassembly of the rotor, which degrades the rotor balance. The bearing sleeves are made of ZrO2 ceramics, and have a diameter of 4 mm and a length of 2.7 mm. Each of Transactions of the ASME of Use: http://www.asme.org/about-asme/terms-of-use b 0 i p e T i a b r 0 b a d s h 1 n t b b F c F j J Downloaded Fr oth thrust and journal bearings has eight gas supply ports of .3 mm diameter. The gas pressure supplied to the journal bearngs can be changed for every two gas supply ports to provide ressure anisotropy. The designed load capacities of the hydroinrtia gas journal tested and thrust bearings are shown in Fig. 6. he design method of the hydroinertia gas bearings was reported n Ref. 10 . The maximum load capacity appears around 25 m nd 20 m of the bearing clearance for the journal and thrust earings, when the eccentricity , which is the displacement of the otor from the center position divided by the bearing clearance, is .5 and 0.6, respectively. The bearing clearance of the journal earing is defined by the difference between the bearing radius nd the shaft radius. The bearing clearance of the thrust bearing is efined by the clearance between the thrust disk and the bearing leeve, when the thrust disk is located at the center of the bearing ousing. In this study, the bearing clearances are 30 m and 7 m for the journal and thrust bearings, respectively. The joural bearing clearance is set slightly larger than the value at which he maximum load capacity is obtained for higher whirl stability, ecause the hydrodynamic effect, which is the cause of whirl, ecomes smaller in a larger bearing clearance. ig. 5 Half-splitting type bearing sleeves made of ZrO2 eramics ig. 6 Load capacity and gas flow rate of the hydroinertia gas ournal \u201ea\u2026 and thrust \u201eb\u2026 bearings ournal of Tribology om: http://tribology.asmedigitalcollection.asme.org/ on 01/29/2016 Terms"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002448_icelmach.2010.5608187-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002448_icelmach.2010.5608187-Figure5-1.png",
+        "caption": "Fig. 5. Operational deflection for 60fmech at 3240 Hz, 3240 r/min, and 40 Nm",
+        "texts": [
+            " The reason for the rotation opposite to the rotor is that the fundamental component of the air-gap field and the 5th spatial harmonic rotate in opposite directions. Fig. 4 shows 40fmech at 3867 Hz, 5800 r/min, and 20 Nm. The deflection shape is similar but rotates clockwise. From observation of both this and the previous figure it becomes apparent that tangential forces also contributes to the surface normal vibrations and therefore the acoustic sensation. Tangential forces can strengthen the bending motion of the stator and may therefore contribute to the surface normal vibration. This will be investigated in the next section. Fig. 5 for 60fmech at 3240 Hz, 3240 r/min, and 40 Nm now shows a significantly different behavior. The deflection shapes of both stator and housing again resemble that of the excitation shape \u2013 but which is mode 0 in this case. The resulting breathing deflection is apparent. The more blueish color indicates a lower surface normal velocity lower. This can also be seen from Fig. 2(a). The literature on electromagnetically excited acoustic noise in electric machines points out the dominance of radial forces [4]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002980_2011-01-0956-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002980_2011-01-0956-Figure1-1.png",
+        "caption": "Figure 1. FMVSS 135 TYPE 2 Compliance Center of Gravity Trapezoid from the Incomplete Vehicle Document [2].",
+        "texts": [
+            " TYPE 2 compliance guidelines benefit the final stage manufacturer compared to TYPE 3 because of the cost associated with developing ESC calibrations and testing for FMVSS 126 compliance. A method to find TYPE 2 compliance guidelines for FMVSS 126 was needed. Past experience with FMVSS 135, Hydraulic Brake Systems, has shown that restricting the center of gravity location of the completed vehicle in the longitudinal and vertical directions is sufficient to achieve TYPE 2 compliance if the components listed in the incomplete vehicle document, including the complete brake hardware, are not modified. Figure 1 shows the trapezoid used to define the FMVSS 135 center of gravity location boundaries. Because the sine with dwell maneuver uses the steering wheel and builds lateral acceleration, the basic concept of restricting the center of gravity location was expanded to include the lateral direction. A combination of physical and HIL testing was used to define TYPE 2 compliance guidelines by restricting the center of gravity location of the completed vehicle. Outside of the scope of this paper, a robust ESC calibration was developed for the incomplete vehicle using ballasted physical vehicles",
+            " FMVSS 301 and FMVSS 135 requirements were used as a starting point for analysis as they place restrictions on mass, vehicle center of gravity location, and the type and size of service body attached to the incomplete vehicle. The pertinent FMVSS 301 requirements found in the incomplete vehicle document are summarized in Table 5 found in the Appendix. Both FMVSS 135 and FMVSS 301 contain explicit center of gravity location boundaries. The center of gravity location boundaries used in the FMVSS 135 trapezoid, shown in Figure 1, are listed in Table 1. Note that the table values are referenced to the front axle centerline and the surface on which the vehicle sits. The FMVSS 135 trapezoid proved to be less restrictive than the final FMVSS 126 requirements, but still proved to be a useful guideline for FMVSS 126 TYPE 2 compliance. The guideline for FMVSS 301 compliance is actually the most restrictive requirement on the vertical direction, but the final-stage manufacturer is able to independently certify FMVSS 301 compliance for higher center of gravity locations if needed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002435_j.triboint.2012.01.011-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002435_j.triboint.2012.01.011-Figure1-1.png",
+        "caption": "Fig. 1. Schematic and settings of tester.",
+        "texts": [
+            " Given the generally larger extent of this kind of leakage, there should be another mechanism of leakage other than those described above. In this study, we used a simple test rig to demonstrate that another significant mechanism of leakage exists for above cases, and to explain the large leakage. We also investigated the leakage phenomena that could occur when the shaft is inclined against the bearing or there is precession motion in the shaft-and-bearing assembly. A simple test rig was used to investigate the phenomenon of grease leakage [14]. Fig. 1(a) shows a schematic of the tester. This tester consisted of a rotating stainless steel disk (JIS SUS 304 stainless steel, 50 mm in diameter) and a glass plate separated by a small gap. In the tester, the circumferential motion of actual bearings was transformed into lateral motion for an easy visualization of grease movement. The disk corresponded to the shield plate of the bearing, and the plate to the inner race. The gap was set to 0.1 mm\u2014the typical gap between the shield plate and inner ring in the shield bearing. Grease was deposited over the disk and the glass plate, forming a right triangular cross-section of 4.5 mm inside as shown in Fig. 1(a). This simulated the situation in a bearing where grease lay between the shield plate and inner race, and adhered to both. This situation is common for actual bearings filled with a sufficient amount of grease. A CCD camera placed under the glass plate and a high speed camera above the grease were used to observe grease movement. The seepage of grease into the gap between the disk and glass plate was equivalent to the flow of grease leaking into the gap between the inner race and shield plate of an actual bearing",
+            " There were apparently no pushing-out forces generated by the rolling elements or cage, and there was no centrifugal force forcibly moving the grease into the gap on the test rig. Centrifugal force only forcibly moved the grease outward from the gap. Airflow in the gap space did not affect grease movement, which was checked using powdered milk instead of grease. Thus, any grease observed seeping into the gap would mean that the mechanisms of grease leakage did not entail centrifugal force or the pushing-out force generated by the rolling elements or cage motion. The disk and glass plate were set in the three positions shown in Fig. 1(a) and (b): the normal position (CASE 1), an inclined disk that leads to disk precession motion (CASE 2), and an inclined glass plate (CASE 3). As shown in Fig. 2, these positions correspond to the correct bearing and shaft positions, shaft inclination against the bearing, and shaft precession motion in the actual bearing-and-shaft assembly, respectively. Tables 1 and 2 list the test greases and test conditions, respectively. No. 3 PFPE grease was applied for CASE 1 and No. 4 for CASES 2 and 3, in considering the assumed actual application of each type of grease"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003048_1.3650825-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003048_1.3650825-Figure5-1.png",
+        "caption": "Fig. 5 Sectional v i ew of the test bearing assembly",
+        "texts": [
+            " As the inlet oil temperature rises, the load capacity is seen to drop, although the film thickness at which a given load is carried remains virtually constant. Fig. 4 shows the effect of speed on the load capacity for a given inlet oil temperature. At 300 deg, the load capacity is seen to be independent of speed, although the film thickness at. which this load is carried varies. Since the load capacity is thermally induced, a given load requires a certain shearing rate, and for that CO load, \u2014 = const. h2 Apparatus and Measurements Apparatus. A sectional view of the test bearing assembly is shown in Fig. 5. The high-speed shaft was 5/s in. dia and carried a 3-in. dia steel thrust disk. The opposing faces of the thrust disk were machined and subsequently lapped so that they D E C E M B E R 1 9 6 5 / 8 2 7 Downloaded From: http://fluidsengineering.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jfega4/27267/ on 06/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use THERMOCOUPLES F I L M T H I C K N E S S PROBE E N T R A N C E 1\u20141/8 \\ i i \u00ae LOAD {THEORY) \u00a9 LOAD ( E X P l ) \u00a9 FRICTION ( E X P l ) \\ \\ \\ \u00a9 \u2014 / o' / \u00b0 \\ \u00b0 / V x o x a \\ - 0",
+            " The essential quantities measured under test were speed, load, torque, inlet oil pressure, temperature, and the film thickness. The rotational speed of the high-speed shaft was measured by a standard tachometer. The hydraulic jacks were loaded by a pressure gage tester which was capable of supplying pressures up to 500 lb/sq in. in increments of 5 Ib/sq in. The friction torque transmitted to the thrust pads was measured by means of a variable reluctance load cell. The static pressure of the oil supply was measured by means of four pressure taps in the annulus of one of the bearing pistons, Fig. 5. The four taps were connected to a common manifold, from which an average static pressure reading could be taken by a Mercury manometer. The temperature of the oil was measured at the inlet (and outlet where applicable) of each sector of the thrust pad at the points shown in Fig. 6. Two copper-constantan thermocouples were placed in each oil groove. The thermocouples were mounted in tufnol inserts which were screwed into the brass backing plate of the thrust pad. The approximate disk temperature was measured by means of the thermistor probe shown in Fig. 5. This 0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 FILM THICKNESS (in X 10\"3 ) Fig. t ies 9 C o m p a r i s o n of theore t ica l a n d e x p e r i m e n t a l bea r ing l oad capac i - a n d expe r imen ta l coef f ic ients of f r i c t ion fo r a f i v e - p a d bea r ing device was moved into contact with the stationary disk at the conclusion of each run. The thickness of the fluid film in the bearing was measured b}r means of a linear displacement transducer of the variable reluctance type. This transducer had a range of \u00b10.040 in. and was fitted with a special steel tip which had a hemispherical end of radius 0.030 in. The transducer was screwed into the loading piston crown, as shown in Fig. 5, and the plunger tip passed through the thrust pad to make contact with the thrust disk. The electrical output of the transducer was therefore a measure of the distance between the thrust pad and the thrust disk. The differential transformer indicator used had a maximum resolution of 10 microin. 8 2 8 / D E C E M B E R 1 9 6 5 Transactions of the AS M E Downloaded From: http://fluidsengineering.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jfega4/27267/ on 06/07/2017 Terms of Use: http://www"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003031_iros.2011.6094957-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003031_iros.2011.6094957-Figure5-1.png",
+        "caption": "Fig. 5. Graspability map for a banana and the DLR hand II: a) Obtained using Algorithm 1; b) Obtained from a database with 20,000 FC grasps; c) Including the grasp quality information.",
+        "texts": [
+            " This section presents some examples of graspability maps using two benchmark objects: a banana and a coffee cup, described with 1,703 and 2,731 points, respectively. The computation of the maps was implemented using Matlab in a commodity desktop computer. For the map creation, the parallelepiped enclosing the space around the object was voxelized with voxels of side length 20 mm. To generate the potential poses of the set \u0393, 16 points were defined on the sphere inside each voxel, with 12 possible orientations per point (i.e. 192 possible poses per voxel). Fig. 5a shows a visualization of the graspability map for the banana and a 4-finger DLR hand II. The computation using Algorithm 1 takes 7.1 hours. The spheres are colored according to the number of frames on the sphere that allow an FC grasp. Blue indicates regions from which the object is well graspable, and red indicates regions from which the object can only be grasped with few hand orientations. This number, however, does not provide information about directional preferences. It can be seen how the ability of a hand to grasp an object changes around the object. Fig. 5b shows the graspability map for the same object and hand, obtained as a graphical representation of a database with 20,000 FC grasps generated with the grasp algorithm presented in [10]. This representation is equivalent to obtaining a graspability map using a sufficient condition for FC, i.e. all the included grasps are FC, but there are more possible FC grasps on the object which are not considered. Of course, Fig. 5b is a subset of Fig. 5a. On the other hand, Fig. 5a might include poses which actually do not lead to FC grasps, due to the employment of a necessary condition (Section III-E). The real graspability map should lie somewhere in between the two presented maps. The advantage of using a database for creating a graspability map is that the maximum grasp quality achievable from a given voxel can be visualized also on the same map, as shown in Fig. 5c; blue indicates regions which lead to a high quality grasp. This information is very useful for selecting the most promising directions for obtaining a very robust grasp. Fig. 6 presents the same comparison for a different object, a coffee cup, using the same hand. The computation takes 13.7 hours in this case. The graspability map is also a useful tool to compare the grasp capabilities of different mechanical hands. For instance, Fig. 7 show the graspability maps for the banana using a 3-fingered Barrett hand, a 4-fingered DLR hand II, and a 5-fingered DLR-HIT hand II [28]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002181_te.2010.2043844-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002181_te.2010.2043844-Figure6-1.png",
+        "caption": "Fig. 6. Motor speed measurement.",
+        "texts": [
+            " The initial unloaded motor exercises were of a more observational than quantitative nature. Students were directed to connect the three-phase motor to their source (with a small-value resistor in series with each lead) and observe the motor\u2019s rotation. The students then interchanged any two motor leads, thereby reversing the phase sequence, and noted any changes in the motor performance (the motor reverses). Students constructed a simple tachometer using a slotted disk attached to the motor shaft and a transmissive optoswitch as a detector [1] as shown in Fig. 6. Motor speed was measured and compared to the frequency of the input three-phase power. The number of motor poles, , was determined by noting that the rotational speed of the motor (in RPM) is related to the frequency of the input electrical power, , by Students were reminded that magnetic field poles come in pairs. If the students were using a variable-frequency source, they were asked to change the frequency of input power source and note that the motor rotational speed varies directly with frequency"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure1.3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure1.3-1.png",
+        "caption": "Figure 1.3 Fluid level comparator",
+        "texts": [
+            " For the most part we shall be dealing with \u2018linear\u2019 systems where the action taken by the control system is in proportion to the magnitude of the difference between the desired value and the actual output. There are many examples of summation/comparison devices used in all engineering disciplines including mechanics, hydraulics, pneumatics and electronics. Presented below are just a few of the more common types. The Float Control Valve Everyone is familiar with the float mechanism inside the toilet cistern that limits the level of water in the cistern during refill following the flush action. Figure 1.3 shows this system both schematically and in block diagram form. This same concept is used extensively in aircraft fuel systems where float valves are used to provide a signal to the refueling or fuel transfer system that the fuel tank is full. In this case the float valves do not act directly on the flow control valve but provide a servo pressure signal that initiates control valve action shutting off flow to that tank. For this reason these valves are termed \u2018float pilot valves\u2019. Mechanical Linkage Summing Mechanical linkage summing is an extremely simple form of summing device that is used extensively in many aircraft flight control systems in service today"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001470_j.jsv.2008.05.001-Figure15-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001470_j.jsv.2008.05.001-Figure15-1.png",
+        "caption": "Fig. 15. Spring\u2013damper model of self-impact in a double pendulum.",
+        "texts": [
+            " The case of plastic impact, w \u00bc 0, is special and characterised by smooth phase-plane trajectories with two impacts that have different flavours. One of the impacts has no sticking phase at all and the separation is immediate even though it is modelled as plastic. ARTICLE IN PRESS S. Singh et al. / Journal of Sound and Vibration 318 (2008) 1180\u20131196 1195 Appendix The method used in this derivation follows closely the method of calculating the coefficient of restitution in terms of spring\u2013damper terms for liner impacts as described in Chatterjee et al. [6]. The mathematical model of impact in the double pendulum is shown in Fig. 15. The contact dynamics is represented by the following set of equations: \u00f0m1 \u00fe 3m2\u00de 3 l21 \u20acy1 \u00fe m2l1l2 \u20acy2 cos\u00f0y1 y2\u00de 2 \u00fe m2l1l2 _y 2 2 sin\u00f0y1 y2\u00de 2 \u00fe \u00f0m1 \u00fe 2m2\u00de 2 gl1 sin y1 \u00fe k\u00f0y1 y2\u00de \u00fe c\u00f0_y1 _y2\u00de \u00bc 0 (24) m2l 2 2 \u20acy2 3 \u00fe m2l1l2 \u20acy1 cos\u00f0y1 y2\u00de 2 m2l1l2 _y 2 1 sin\u00f0y1 y2\u00de 2 \u00fe m2gl2 sin y2 2 \u00fe k\u00f0y2 y1\u00de \u00fe c\u00f0_y2 _y1\u00de \u00bc 0 (25) In the presence of high impact force the gravity force terms can be neglected. Also, in the impact zone: y1 y2E0, giving the following reduced equations: \u00f0m1 \u00fe 3m2\u00de 3 l21 \u20acy1 \u00fe m2l1l2 \u20acy2 2 \u00fe k\u00f0y1 y2\u00de \u00fe c\u00f0_y1 _y2\u00de \u00bc 0 (26) m2l 2 2 \u20acy2 3 \u00fe m2l1l2 \u20acy1 2 \u00fe k\u00f0y2 y1\u00de \u00fe c\u00f0_y2 _y1\u00de \u00bc 0 (27) Putting f \u00bc y1 y2, the equations are further reduced to I \u20acf\u00fe c _f\u00fe kf \u00bc 0 (28) where I \u00bc \u00f04rm \u00fe 3\u00dem2l21 6\u00f06r2l \u00fe 6rl \u00fe 2rmr2l \u00fe 2\u00de (29) rm \u00bc m1 m2 ; rl \u00bc l1 l2 At the beginning of the contact phase, t \u00bc 0, f \u00bc 0, df/dt \u00bc o"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000000_acc.2006.1657181-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000000_acc.2006.1657181-Figure1-1.png",
+        "caption": "Fig. 1. PMSM circuit model with stator winding short",
+        "texts": [
+            " PMSM dynamics under both healthy and faulty operations can be simulated with lumped parameter models based on coupled magnetic circuit principle [15]. Modeling of the healthy PMSM with balanced three-phase input has been widely recognized, which are generally expressed in dq-axis rotor reference frame by applying Park transformation [15][16]. This section focuses on the introduction of PMSM modeling under stator interturn short in one of its phase windings. This model directly serves the fault diagnosis purposes. Fig. 1 shows the faulty PMSM magnetic circuit model, where a partial interturn short is illustrated in the stator winding of phase b. In order to take into account the presence of the shorted circuit in the PMSM model, we need to partition the affected phase windings and include an extra circuit for that phase. This extra circuit creates a stationary magnetic field, which modifies the original field by adding the fourth coupled magnetic circuit into the system. To clearly represent the fault size and its location, two new parameters, \u03c3 and \u03b8f, are introduced"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001647_9780470630693.ch5-Figure5.27-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001647_9780470630693.ch5-Figure5.27-1.png",
+        "caption": "FIGURE 5.27 A scheme of self-assembly of thiol functionalized and bilirubine oxidase appended MWCNTs on gold electrode [92].",
+        "texts": [
+            " The 202 BIOCATHODES FOR DIOXYGEN REDUCTION IN BIOFUEL CELLS macroporous structure was synthesized electrochemically by the inverted colloidal crystal template technique. This allows one to control the thickness of the film and the number of nanoparticles and protein molecules. Although the onset potential of catalytic current remained almost unchanged in comparison with the smooth electrode, it was applied to the dioxygen\u2013glucose biofuel cell [153]. The most recent example of bioelectrocatalysis on gold electrodes employs the affinity of thiol-modified MWCNTs with covalently bonded bilirubin oxidase [92] (Fig. 5.27). The design of a biocathode, as of any enzyme-modified electrode, is not an easy task because of several tough requirements to be fulfilled. They include selection of appropriate enzyme and (in some cases) the mediator to provide a high potential of oxidant bioelectrocatalytic reduction. Its high efficiency depends on the same factors and electrode construction. At the same time, a friendly environment for fragile protein immobilization has to be provided. This is crucial from the practical perspective because stable performance of the device is expected"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003972_icra.2011.5980334-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003972_icra.2011.5980334-Figure6-1.png",
+        "caption": "Fig. 6. Uncertainty in the contact location results in a displacement of the hyperplanes defining the search zones Si.",
+        "texts": [
+            " The primitive wrenches produced at the potential locations of the real contact point are described with !ij = ( n\u0302ij pib \u00d7 n\u0302ij ) = ( n\u0302ij pi \u00d7 n\u0302ij ) + ( 0 \u0394pi \u00d7 n\u0302ij ) (8) Thus, the uncertainty in the location of the contact point is a perturbation \u0394 affecting only the torque components of the wrench. Note that the magnitude of n\u0302ij in Eq. (8) is 1, so the magnitude of the maximum perturbation in the torque direction is \u2225\u0394 \u2225max = \u2225\u0394pi \u00d7 n\u0302ij\u2225 = \u2225\u0394pi\u2225 (9) To illustrate the effect of this perturbation in the computation of the ICRS, Fig. 6 illustrates a hypothetical 2-dimensional wrench space, with the horizontal axis representing the force component f and the vertical axis representing the torque component for the wrench. Let a generic hyperplane Hk be described with the equation e \u22c5 ! = e0, where e is the vector normal to the hyperplane. The distance of the hyperplane to the origin is given by D = \u2223e0\u2223 \u2225e\u2225 (10) Now, let every point of a hyperplane H \u2032\u2032 k (which partially defines a search zone Si in Algorithm 1, Section II-C) be moved by a distance \u0394 in the torque direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure3-1.png",
+        "caption": "Fig. 3 The McGill SMG",
+        "texts": [
+            " This is a robot with its four motors located at the base and coupled to the moving plate by means of four limbs, the moving plate being capable of a full rotation by virtue of a special mechanism added at the plate. A simpler architecture, with only two limbs, was prototyped at McGill University\u2019s Centre for Intelligent Machines, as Proc. IMechE Vol. 223 Part I: J. Systems and Control Engineering JSCE623 F IMechE 2009 at Purdue University on June 28, 2015pii.sagepub.comDownloaded from illustrated in Fig. 3. In this architecture, the four motors are located at the base plate, and laid out in two pairs, each pair driving one limb. The limb pair of motors produces both a pan and a tilt motion of the proximal P joints. One distal P joint on each limb is coupled to its proximal counterpart in such a way that the planes of motion of the two limb P joints generate the T 2 subgroup of displacements of the moving plate. The two revolute joints, one coupling the limb to the base and one to the moving plate, complete the four-joint kinematic chain of each limb"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001257_j.euromechsol.2008.11.006-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001257_j.euromechsol.2008.11.006-Figure3-1.png",
+        "caption": "Fig. 3. Cross section of geomembrane tube.",
+        "texts": [
+            " (4a), (4b) may be linearized and combined to give \u22022u \u2202t2 = ( A \u03c1 dP dA ) \u22022u \u2202z2 (5) from which one readily obtains that the velocity of very long waves to be C0 = ( A \u03c1 dP dA )1/2 . (6) The wave propagation velocity can easily be obtained by using Eq. (6). In which the subscript 0 is the long wave limit. If one considers the area to be the equivalent of density in compressible flow, one can appreciate the analogy to compressible flow, since then C0 = (dP/d\u03c1)1/2, the velocity of sound waves (Paidoussis, 2004). Fig. 3 shows the cross section of an inflated geomembrane tube resting on rigid foundation. Points A and B are right and left contact points between the tube and the rigid foundation. The coordinates are X(S) and Y (S), and the angle of the tangent with the horizontal direction is \u03c8(S), where S is the arc length and measured from point A(S = 0). The total perimeter is L, the contact length is C and the width of the tube is assumed unity. In the next subsections, the equilibrium shapes of a geomembrane tube are obtained for three different internal pressures"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002683_j.finel.2010.08.001-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002683_j.finel.2010.08.001-Figure11-1.png",
+        "caption": "Fig. 11. Right angle frame: geometry and loading definitions.",
+        "texts": [
+            " As a predictor point u(k) (0) we choose the linear extrapolation of the previously computed u(k 1) and u(k 2) vectors when k41, while the formula u\u00f01\u00de \u00f00\u00de \u00bc u\u00f00\u00de \u00feDt\u00f01\u00de _u\u00f00\u00de is used when k\u00bc1. Once setted an initial Dt\u00f01\u00de value, so that stable solutions are obtained, also here the adapted based length (28) of the step has been used to evaluate the predictor points Dt\u00f0k\u00de,u\u00f0k\u00de \u00f00\u00de. A set of examples is examined to illustrate the features of the presented formulation. In particular, the tests analyse plane and spatial kinematics by modelling the body with the described twoand three-dimensional elements. A right angled frame shown in Fig. 11, fully restrained at one end, is analysed both in statical and dynamical cases. Equilibrium states were computed by the three-dimensional finite element formulation. It is worth noting that the motion of the system involves large torsion and bending. configurations at marked solution points. namical solution curves. In the statical case, the analysed frame is loaded at the free end B by f u B \u00bc l in the x1 direction. An imperfection load f w B \u00bc 10 4l in the x3 direction is set to initiate lateral buckling along the fundamental equilibrium path"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003761_s00170-013-4988-8-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003761_s00170-013-4988-8-Figure8-1.png",
+        "caption": "Fig. 8 Temperature fields of the die casting in the solidification stage (t=4.93 s). a Temperature field of outer surface. b Temperature field of inner surface",
+        "texts": [
+            "23 s for the second plan, while the first one still needs some more time to complete the filling process. In addition, this optimization in fluid flow condition will lead to more smooth temperature field as the liquid metal could separately flow into two sides at almost the same time. So, the second plan was adopted for the new die-casting die. The cooling stage is simulated. High-temperature regions of the die-casting part and the die during the cooling stage for the second plan are shown in Figs. 8 and 9 separately. Figure 8 shows that high-temperature regions are centralized on the ribs with small radius in Fig. 8a and the central platforms in Fig. 8b. Figure 9 shows the temperature of the die halves. The high-temperature regions are on the ribs and the central platforms as well, corresponding to the defect regions of die-casting product. As mentioned above, hightemperature regions are prone to generate thermal cracks during quenching, and the edges of these regions are the weakest parts. So, it explains why most of the cracks on the surface of experimental die-casting product and mold were concentrated on the edges around the small radius of the ribs and the boundary edges of the platforms (as shown in Figs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002377_elan.200900413-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002377_elan.200900413-Figure7-1.png",
+        "caption": "Fig. 7. Schematic representation for the proposed biosensor.",
+        "texts": [
+            " Figure 6B shows the calibration curve of the GOD/PAH/MnOx electrode response to glucose, based on data in Figure 6A. The anodic currents were in proportion to the concentration of glucose from 19.6 to 107.1 mM. The fitted linear equation is y\u00bc 0.2607 x\u00fe 7.642 (R2\u00bc 0.9862), where y and x stand for the peak current (mA) and the concentration (mM) of glucose, respectively. The limit of detection was estimated to be 13.0 mM at a signal/noise ratio of 3. The mechanism of the liposomal GOD/PAH/MnOx electrode for the detection of glucose can be represented schematically in Figure 7. Glucose and O2 diffuse from the bulk solution into the liposome, and then glucose is enzymatically oxidized by the liposome-encapsulated GOD to produce gluconolactone and H2O2. The liposome membrane allows the penetration of the generated H2O2, followed by the catalytic oxidation of H2O2 at active sites (Mn4\u00fe) in the manganese oxide. As soon as the Mn sites are reduced to the lower state (Mn3\u00fe), they are oxidized back to Mn4\u00fe via electron exchange on the electrode surface. As a result, when the series of reactions repeats continuously, an electric current flow occurs proportionally with the glucose concentration"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-7-1.png",
+        "caption": "Figure 3-7: Normal-load zone, outer ring rotating relative to load, or load rotating in phase with inner ring.",
+        "texts": [
+            " Figure 3-5 illustrates how an applied load of constant direction is distributed among the rolling elements of a bearing. The large arrow indicates the applied load, and the series of small arrows show the share of this load that is supported by each ball or roller in the bearing. The rotating ring will have a continuous 360 zone, while the stationary ring will show a pattern of less than 180 . Figure 3-6 illustrates the load zone found inside a ball bearing when the inner ring rotates and the load has a constant direction. Figure 3-7 illustrates the load zone resulting if the outer zone rotates relative to a load of constant direction, or when the inner ring rotates and the load also rotates in phase with the shaft. Figure 3-8 illustrates the pattern found in a deep-groove ball bearing carrying an axial load, and Figure 3-9 shows the pattern from excessive axial load. Uniformly applied axial load and overload are the two conditions where the load paths are the full 360 of both rings. Combined thrust and radial load will produce a pattern somewhere between the two, as shown in Figure 3-10"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000742_1.2736732-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000742_1.2736732-Figure2-1.png",
+        "caption": "Fig. 2 \u201ea\u2026 3D pressure distribution in the slider bearin the centerline \u201ez=1/2\u2026 of the slider bearing for Bingha",
+        "texts": [
+            " 4 Results and Discussion Different types of calculations have been done for two typical lubrication problems: a slider bearing and channel flow parallel plates . In order to compare our results with the literature, the properties of lubricant are taken as follows Cp = 1966 J/kg \u00b0 C, Kc = 0.1 W/m \u00b0 C, T0 = 40 \u00b0 C, and \u0304 = 870 kg/m3 4.1 Slider Bearing Flow. We consider the three-dimensional thermohydrodynamic flow in a slider bearing with H\u0304=0.1 mm, L= B\u0304=100 mm, and U\u0304=3.6 m/s. First, we have plotted the three dimensional 3D pressure distribution in the slider bearing, using Bingham lubricant film Fig. 2 a . Our results are in close agreement with Kim and Seireg 3 for a parameter m equals to 9.72 10\u22125 Fig. 2 b . They have validated their computational results by experimental ones 11 . In continuation of our computational study, we use the nonNewtonian film parameters listed in Table 1. In the Bingham and the Herschel\u2013Bulkley models, the parameter m is set equals to 3.6 107. This selected value allows an excellent similitude between the modified constitutive equations of Papanastasiou Eq. 8 and the classical law available in the literature Fig. 3 . For an inlet\u2013outlet film height ratio a equal to 2 and a transverse incline ratio b equal to 1, Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002212_lars.2010.27-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002212_lars.2010.27-Figure1-1.png",
+        "caption": "Fig. 1. Mobile robot geometry.",
+        "texts": [
+            " The Pioneer P3-AT robotic platform used in this work can be modeled as a four-wheel skid-steering vehicle actuated by two motors, one to drive the right-sided wheels and another to drive the left-sided wheels. In this section the kinematic and dynamic models of the skid-steering mobile robot (SSMR) based on [3], [4] are presented. These models are used by the motion planning method, in Algorithm 2 line 5, to compute the future state of the robot given its current state and control inputs. The geometry of the robot is presented in Fig. 1, where (X,Y ) is the inertial reference frame, (X \u2032 , Y \u2032 ) the reference frame fixed on the robot with origin on its center of mass CG. The position of CG with respect to the inertial frame is (x, y), and \u03c8 is the orientation of the frame fixed on the robot w.r.t. the inertial frame (i.e., the angle between theX \u2032 andX axis). The coordinates of the instantaneous center of rotation (CIR) are (x \u2032 CIR, y \u2032 CIR) in the robot\u2019s frame. Also, a is the distance between the center of mass and the front wheels axle along X \u2032 , b is the distance between the center of mass and the rear wheels axle along X \u2032 , c is half the distance between wheels along Y \u2032 , and r is the wheel radius",
+            " On a curved trajectory, CIR is at a defined x \u2032 CIR along X \u2032 . However, x \u2032 CIR must be between the robot\u2019s two axles, otherwise the vehicle skids drastically with lost of movement stability. Thus, to complete the kinematic model of the SSMR, the following nonholonomic constraint is introduced [3]: \u03c5y\u2032 = x \u2032 CIR\u03c8\u0307, with x \u2032 CIR \u2208 (\u2212b, a). (3) But from (1) we obtain: \u03c5y\u2032 = \u2212 sin\u03c8x\u0307+ cos\u03c8y\u0307. (4) Replacing equation (4) on (3) and writing in the Pfaffian\u2019s form, we have [ \u2212 sin\u03c8 cos\u03c8 \u2212x\u2032 CIR ] \u23a1\u23a3 x\u0307 y\u0307 \u03c8\u0307 \u23a4 \u23a6 = A(q)q\u0307 = 0 From Fig. 1, the velocities q\u0307 can be represented as q\u0307 = S(q)\u03b7, (5) where S(q) = \u23a1 \u23a3 cos\u03c8 \u2212x\u2032 CIR sin\u03c8 sin\u03c8 x \u2032 CIR cos\u03c8 0 1 \u23a4 \u23a6 , \u03b7 = [ \u03c5x\u2032 \u03c9 ] . S(q) is a full rank matrix, whose columns are in the null space of A(q), i.e. ST (q)AT (q) = 0 It is interesting to note that because dim(\u03b7) = 2 < dim(q) = 3, the equation (5) describes the kinematic of a sub-actuated robot with the nonholonomic constraint given by (3). The forces and moments acting on the robot are presented in Fig. 2, where Fx\u2032 i and Rx\u2032 i are the traction forces and the longitudinal resistive forces acting on each wheel"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003246_esda2010-24367-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003246_esda2010-24367-Figure1-1.png",
+        "caption": "FIGURE 1. MISALIGNED CYLINDRICAL ROLLER WITH DUB-OFF PROFILE.",
+        "texts": [
+            "(1) becomes 3 3 ( )12H P H P HU X X Y Y X r r r h h \u00e6 \u00f6 \u00e6 \u00f6\u00b6 \u00b6 \u00b6 \u00b6 \u00b6 + =\u00e7 \u00f7 \u00e7 \u00f7\u00b6 \u00b6 \u00b6 \u00b6 \u00b6\u00e8 \u00f8 \u00e8 \u00f8 (3) The density and viscosity of lubricant can be written in dimensionless form as follows [4]: 0.581 1 1.68 EP EP r = + + (4) exp( )GPh = (5) where G Ea= and E is given in GPa. The EHL film thickness can be written as ( , ) ( , ) ( , )o g dh x y h h x y h x y= + + (6) where oh is a constant, ( , )gh x y is the separation due to the geometry and tilting of the roller in their undeformed state, and ( , )dh x y is the amount of elastic deformation. For a dub-off profiled cylindrical roller under misaligned state as shown in Fig.1, ( , )gh x y is expressed as: 22 ( ) ( , ) tan + 2 2 d g d y yxh x y y R R q - = + (7) where roller misaligned angle q is measured from left end of the roller in the y direction. In the un-profiled regions, the last term should be omitted. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/27/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 3 Copyright \u00a9 2010 by ASME Assuming the contacting rollers and races and races behave as elastic half spaces, the elastic deformation at ( , )x y owing to the pressure distributed over a rectangular element, as shown in Fig",
+            " a footprint half width of aligned roller at x=0 Di,j,k,l dimensionless elastic influence coefficient, , , , /i j k lD R , , ,i j k lD elastic influence coefficient E modified modulus of elasticity, 2 2 1 1 2 22 / (1 ) / (1 ) /E v E v E= - + - G dimensionless materials parameter, Ea h film thickness H dimensionless film thickness, /h R Hm dimensionless absolute minimum film thickness, See Fig.3 Hc dimensionless local minimum film thickness at the middle of roller, See Fig.3 L roller length p pressure P dimensionless pressure, /p E R roller radius Rd dub-off radius, See Fig.1 * dR /dR R u mean entraining velocity in the x direction U dimensionless speed parameter, /ou ERh w normal applied load W dimensionless load parameter, 2/w ER x, y Cartesian coordinates dy See Fig.1 Yd /dy R X, Y / , /X x R Y y R= = d elastic deformation caused by a rectangular pressure element (eq. 8) h dimensionless viscosity, / oh h oh viscosity at atmospheric pressure h viscosity r dimensionless lubricant density, / or r or lubricant density at atmospheric pressure r lubricant density q roller misalignment angle, See Fig.1 [1] Heydari, M., and Gohar, R., 1979, \"The Influence of Axial Profile on Pressure Distribution in Radially Loaded Rollers\", J. of Mech. Eng. Sci., 21, pp. 381-388. [2] Rahnejat, H., and Gohar, R., 1979, \"Design of Profiled Tapered Roller Bearing\", Tribology Int., 12, pp. 269-275. [3] Skurka, J. C., 1970, \"Elastohydrodynamic Lubrication of Roller Bearings\", J. of Lub. Tech., 92, pp. 281-291. [4] Dowson, D., and Higginson, G. R., 1959, \"A Numerical Solution to the Elastohydrodynamic Problem\", J. of Mech"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001806_analsci.26.417-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001806_analsci.26.417-Figure1-1.png",
+        "caption": "Fig. 1 Schematic structure of a fabricated microfluidic polymer chip consisting of a mixing chip and a NO3 \u2013-ISE detector chip. a) Sample solution, b) carrier solution, c) solutions from the mixing chip, d) NO3 \u2013-ISE, e) Na+-ISE used as a reference electrode.",
+        "texts": [
+            " We adopted emf values of which the variation with time reached 0.5 mV/min. In general, a stable emf was obtained within 5 min, and the variation of emf with time was very small. The reproducibility of the potential response was checked by the standard deviation of the average slope values obtained with multiple calibration (n = 3). Microfluidic\u202fpolymer\u202fchip\u202f integrated\u202fwith\u202fa\u202fNO3 \u2013-ISE\u202fdetector\u202f and\u202fflow\u202fsystem\u202ffor\u202fmeasuring\u202fnitrate\u202fion The microfluidic polymer chip is composed of a mixing chip and an ISE detector chip based on a NO3 \u2013-ISE, as shown in Fig. 1. The mixing chip consisted of a sample reservoir, a sample injection channel (a), a carrier solution channel (b) and a mixing channel. Channels of the polymer-based microfluidic chip were fabricated using polystyrene plates (ITEM 70128 500, Tamiya Co. Ltd., Japan) and stainless-steel wires (SUS304-W1, Waki Industrial Co. Ltd., Japan) according to almost the same method.41,53\u201355 The chip design of a polymer-based microfluidic chip detector with an embedded NO3 \u2013-ISE and a Na+-ISE as a reference electrode was similar to that shown in a previous paper"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000328_jmes_jour_1974_016_013_02-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000328_jmes_jour_1974_016_013_02-Figure6-1.png",
+        "caption": "Fig. 6. Forces and bending moments pertinent to an element",
+        "texts": [
+            " (10) where 6, is the skin depth in the workpiece defined by . . . . . . . . . . . . (1 1) This analysis holds for tubes whose wall thickness is larger than 6,. In the case of a steel tube at low discharge frequencies, 6, is large and the magnetic field diffuses readily through the tube-wall so that a pressure correction factor has to be included (14) in equation (8). The shell has a dominant radial motion plus small flexures whose amplitudes increase with time. A typical element of the tube-wall, subject to all the pertinent forces is shown in Fig. 6. It is assumed that flexural perturbations do not influence the magnetic coupling between the tube and the coil, which permits us to consider the motion of each mode independently. Radial unperturbed motion The equation of pure radial motion, ignoring bending, is (12) . . . . . . . . . . . Ne - P + -- rhwo a where No is the hoop force, given by j h12 oedz. -hi2 Closed-form solutions to equations (1)-( 12) are difficult to obtain and numerical solutions are developed for the changes of i, wo and hence P, wo",
+            " Hence Journal Mechanical Engineering Science OIMechE 1974 Vol16 No 2 1974 at UNIV OF PITTSBURGH on March 4, 2016jms.sagepub.comDownloaded from Now, integrating (31) twice, we obtain and . . . . . . . . . . wda = 7(2 - r/rf) (34) where (35) r=- . . . . . . . . . . . . . vo t 2a and It is interesting to note that when r = rf, (34) gives ww/a = 71, which shows that rf represents the final hoop strain and is the same as that given by (30). Buckled (perturbed) motion We now assume a radial perturbation, w, in the radial displacement wo of the deforming element, as shown in Fig. 6. The total radial displacement at time t is (wo+w) where w 4 wo. Under pressure P, shear force S, hoop force Ne and bending moment Me, the tube element simultaneously rotates and moves radially. Assuming No to be independent of 8, and that rotary inertia has negligible influence on the rotation, and taking moments about any point on the element, we have (37) s=+ aMe ax . . . . . . . . . . . . . where SX represents the circumferential length of the element at that moment It is related to the angle 0 which it would subtend, had there been no perturbation, through (38) -- ax-a ae ",
+            " The magnetic loading technique is ideally suited for any future study of buckling tubes because it is easily controllable and the pressure can be applied instantaneously over the whole length of the tubes - features which are not easily obtained by detonation. The author would like to express his gratitude to Prof. W. Johnson for his valuable comments and to Dr. S. Reid for checking through the manuscript. APPENDIX 1 DETERMINATION OF MEMBRANE FORCES AND BENDING MOMENTS IN TERMS OF RADIAL DISPLACEMENT FOR PLASTIC CYLINDRICAL SHELLS In order to solve equation (44) we must express Me and Ne in terms of the radial displacements wo and w. It can be shown from Fig. 6, that the membrane forces and bending moments are given by hi2 \u2019 M x = - . . . . . . . . The stress, and hence the strain-distributions, across the wall-thicknesses are obtained by considering the incompressibility condition, the Levy-Mises flow rule and the constitutive equation. The algebra given below follows closely that of Vaughan and Florence (10) and is included here for completeness. The incompressibility condition requires . . . . . . . . . . . ix + \u20ac6 + iz = o (699) where f,, i e and GZ are the longitudinal, circumferential and radial strain rates"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000318_isorc.2006.1-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000318_isorc.2006.1-Figure3-1.png",
+        "caption": "Figure 3. Multi-actor/multi-sensor (MAMS) model",
+        "texts": [],
+        "surrounding_texts": [
+            "An actor performs appropriate actions through some actuation devices in an event area on receipt of sensed QoE values from sensors. An actuation device is modeled to be an object in this paper. An action is modeled to be the execution of a method on an object. On receipt of a method issued by an actor, the method is performed on an actuation device object. An object may be composed of subobjects. For example, a robot arm is composed of multiple parts like motors and servo controllers [14]. Thus, an object is hierarchically structured in a part-of relation [9]. In a method issued to an object, methods on the subobjects are furthermore invoked. Thus, methods are invoked in a nested manner [6]. Let op(s) show the result obtained by performing a method op on state s of an object in the world. Here, let op1 and op2 be a pair of methods of an object o. Let op1 \u25e6 op2 show a serial execution of methods op1 and op2 on an object o. Let op1 \u2016 op2 denote a parallel execution of methods op1 and op2 on an object o. A method op1 is referred to as equivalent with another method op2 (op1 \u2261 op2) iff the result obtained by performing the method op1 on every state of the object o is the same as op2. \u03c6 shows a null method which does nothing on the object o. There are following relations among a pair of methods op1 and op2 of an object o: 1. op1 and op2 conflict with one another if and only if (iff) op1 \u25e6 op2(s) = op2 \u25e6 op1(s) for some state s, i.e. the result obtained by performing the methods op1 and op2 depends on the computation order. op1 \u25e6 op2 \u2261 op2 \u25e6 op1. 2. op1 and op2 are compatible (do not conflict) with one another iff op1 \u25e6 op2(s) = op2 \u25e6 op1(s) for every state s, i.e. op1 \u25e6 op2 \u2261 op2 \u25e6 op1. 3. op2 absorbs op1 iff op1\u25e6op2(s) = op2(s) for every state s, i.e. op1 \u25e6 op2 \u2261 op2. 4. op2 compensates op1 iff op1\u25e6op2(s) = s for every state s, i.e. op1 \u25e6 op2 \u2261 \u03c6. Here, op2 is referred to as compensating method of op1, denoted by o\u0303p1. 5. op1 is idempotent if op1(s) = op1 \u25e6 op1(s) for every state s, i.e. op1 \u2261 op1 \u25e6 op1. Suppose an air-conditioner object ac supports five methods on, off, up, down, and temp. The air-conditioner object Proceedings of the Ninth IEEE International Symposium on Object and Component-Oriented Real-Time Distributed Computing 0-7695-2561-X/06 $20.00 \u00a9 2006 IEEE ac is turned on, off, up, and down by the methods on, off, up, and down, respectively. In addition, the current temperature is obtained by a temp method. The method temp conflicts with all the other methods. A pair of the methods on and off conflict with the other methods up, down, and temp while a pair of the methods up and down are compatible with one another. The method off absorbs the other methods temp, on, up, and down. The methods on, off, and temp are idempotent. The method up can be compensated by the method down and vice versa, i.e. up = \u02dcdown and down = u\u0303p. The method off compensates the method on. The method off does not always compensate the method on. The effect done by a method op can be removed by performing a compensating method o\u0303p on an object o. Garcia-Molina et al. discuss Saga using the compensation of methods [8]. Yasuzawa et al. [23] discuss unresolvable deadlocks to occur by performing compensating methods in object-based systems where objects are hierarchically structured. In database systems, a transaction can be compensated by some transaction. However, there are uncompensatable methods on an actuation device object. For example, a method print-out cannot be compensated since physical papers are used out in a printer object prt. If a missile is launched, it cannot be compensated. An actor may issue a sequence of methods to objects. For example, a door object is required to be locked after closed. Thus, some multiple methods are required to be sequentially performed in a specified sequence. Suppose a key-lock method is issued after a door is closed. Here, suppose the door cannot be locked due to some trouble. In one case, it is considered to be all right because the door is closed. In another case, the closed door should be locked. Here, all the actions performed have to be undone. That is, the door object is opened again. In the later case, a sequence of close and lock methods should be atomically performed. A sequence of methods to be atomically performed is referred to as transaction [10]. Each method is considered to be a transaction in the former case. A transaction is required to abort if some method in the transaction cannot be successfully performed. Suppose a transaction is required to be undone after performing a method up on the air-conditioner object ac. In order to undo the transaction, a compensating method down of the method up is performed on the air-conditioner object ac, i.e. down = u\u0303p [8, 23]. On the other hand, a method like print-out cannot be compensated. Hence, transactions cannot issue uncompenasatable methods. There are types of actuation device objects with respect to how many methods can be concurrently performed, sequential and concurrent ones. In a sequential device object, at most one method can be performed at a time. For example, a device with a single actuator is a sequential object. On the other hand, a device object is referred to as concur- rent object if multiple methods can be in parallel performed on the object. A robot is an example of concurrent object since arms and vehicles in parallel move. Another point of actions on a device object is that each execution of some method is not assumed to be atomic. It takes a longer time to perform a method on an actuation device object compared with the computation speed of an actor. For example, it takes some seconds to print-out papers on an printer device object while a transaction can be performed in a computer for several seconds. Hence, some method is not atomically performed, i.e. the execution of a method may be interrupted. An atomic method should be compensated if the method is partially performed due to some fault."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002743_icmtma.2010.383-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002743_icmtma.2010.383-Figure1-1.png",
+        "caption": "Figure 1. Two-dimensional homing guidance geometry",
+        "texts": [
+            " The paper is organized as follows: in Section 2 the mathematic model of the integrated guidance and control for homing missile against ground fixed target with impact angular constraint, according to the dynamic of autopilot and the nonlinear relative motion between homing missile and target in terminal guidance phase. In Section 3 the nonlinear robust control law is presented based on sliding mode control theory. Section 4 presents numerical simulation results and Section 5 summarizes the final conclusions. II. MODEL OF ENGAGEMENT The section presents the mathematic model derivation of the integrated guidance and control system for homing missile against ground fixed target in the pitch plane. As shown in Fig.1, in the pitch plane dynamics of the homing missile against ground fixed target range and the line-of-sight angle are governed by 978-0-7695-3962-1/10 $26.00 \u00a9 2010 IEEE DOI 10.1109/ICMTMA.2010.383 480 cos( ) sin( ) r V q rq V q \u03b8 \u03b8 = \u2212 + = + (1) where , ,r q \u03b8 are the relative distance, the line-of-sight angle and flight path angle of homing missile. Linearized missile dynamic in the pitch plane is described by [9] 4 1 2 3 z z z z w w a w a w a a \u03b1\u03b1 \u03b1 \u03b1 \u03b4 = \u2212 + \u0394 = + + + \u0394 (2) where \u03b1 is the attack angle, zw is the angular pitch rate, z\u03b4 is the deflection angle for the pitch control, 1 2 3 4, , ,a a a a are the aerodynamic parameters of homing missile, \u03b1\u0394 and w\u0394 are unknown bounded uncertainties"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000021_esda2006-95565-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000021_esda2006-95565-Figure5-1.png",
+        "caption": "FIGURE 5. CUTTING AND FLICKING BRUSHES",
+        "texts": [
+            " The bristles of gutter brushes form a cup-like shape and are slender pieces of material (e.g., steel or polypropylene) that are grouped into clusters, which are arranged in one or more rows. Figure 4 illustrates a gutter brush along with the geometric parameters: bristle length, l, mount angle, \u03c6, and mount radius, r1j, where j is the cluster row number (in the figure, there is only one row of clusters). There are two types of gutter brushes in the market, cutting and flicking brush [14], which differ only in the bristle mount orientation angle, \u03b3, as shown in Fig. 5. In a cutting brush (\u03b3 = 0), the bristle tends to deflect mainly in the radial direction, \u201ccutting\u201d through debris. In contrast, a flicking brush bristle (\u03b3 = 90\u00b0) mainly deflects in the tangential direction, \u201cflicking\u201d debris when it is released from the road at relatively high speed. 3 Copyright \u00a9 2006 by ASME rms of Use: http://www.asme.org/about-asme/terms-of-use Do Apart from the material properties and geometric parameters, the performance of the brush depends on the operating parameters. Some of them are the brush offset angle, \u03be, brush angle of attack, \u03b2, brush rotational speed, \u03c9, vehicle speed, v, brush penetration, \u2206, brush vertical force, Fb, and brush torque, T, which are illustrated in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002181_te.2010.2043844-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002181_te.2010.2043844-Figure3-1.png",
+        "caption": "Fig. 3. Fabricated synchronous motor mount.",
+        "texts": [
+            " In order to make appropriate measurements for motor performance, three items need to be fabricated: a slotted disk (used, in conjunction with a transmissive optoswitch, to form a simple tachometer for motor speed measurements), a motor mount, and a small-scale dynamometer. The slotted disk was cut on the department\u2019s laser cutter and glued onto the motor\u2019s concentric, rotating housing. The motor mount was fabricated from a U-shaped piece of flat plastic, also cut on the laser cutter, and two DIP sockets, used for mechanical mounting of the motor assembly on the protoboard while serving no electrical function. This mount is shown in Fig. 3 with the synchronous motor and the motor\u2019s original mount screwed in place. Care was taken with alignment so that the DIP sockets could be inserted into a typical electrical protoboard and the tachometer\u2019s slotted disk would operate without interference. The motor small-scale dynamometer was based on the spring scale-clothes pin dynamometer used in a dc motor experiment at USD [1] and is described in more detail in the next section. The experimental procedure was broken into four distinct segments: two of which (Y-connected loads and -connected loads) dealt with basic balanced three-phase systems and two (unloaded motor operation and loaded motor operation) with synchronous motor characteristics"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure13-1.png",
+        "caption": "Figure 13. Selected mode shapes of the extended model of transmission, propeller shaft and rear axle",
+        "texts": [
+            " This means that the approach can be used for virtual prototyping where absolute vibration levels are required for NVH target setting purposes. A feasibility study was then performed to extend the modelling to include the drive shaft and the rear axle, as well as the transmission. This model included the driveshafts, joints, and the hypoid gear and internal bevel gears. The extended model is shown in Figure 12. The mode shapes of the system were predicted for a specific load condition. These are shown in Figure 13. The response of the whole system to the planetary transmission error is shown in Figure 14 for the torsional response on the propshaft, and in Figure 15 for the vibration response on axle differential case and on the transmission nose. There is currently no test data available for comparison with these predictions. PARAMETRIC INVESTIGATION It is often interesting and useful to investigate variants of a baseline analytical model. For example, an engineer may want to study the functional effects of the variability of various parameters of a design due to manufacturing tolerances of clearances or gear micro-geometry profiles"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.27-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.27-1.png",
+        "caption": "Fig. 3.27. Velocity and force manipulability ellipses for a 3-link planar arm in a typical configuration for a task of actuating force and velocity",
+        "texts": [
+            "26 the typical configuration of the human arm when writing can be recognized. An opposite example to the previous one is that of the human arm when throwing a weight in the horizontal direction. In fact, now it is necessary to actuate a large vertical force (to sustain the weight) and a large horizontal velocity (to throw the load for a considerable distance). Unlike the above, the force (velocity) manipulability ellipse tends to be oriented vertically (horizontally) to successfully execute the task. The relative configuration in Fig. 3.27 is representative of the typical attitude of the human arm when, for instance, releasing the ball in a bowling game. In the above two examples, it is worth pointing out that the presence of a two-dimensional operational space is certainly advantageous to try reconfiguring the structure in the best configuration compatible with the given task. In fact, the transformation ratios defined in (3.127) and (3.128) are scalar functions of the manipulator configurations that can be optimized locally according to the technique for exploiting redundant DOFs previously illustrated"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003749_s1068366612020110-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003749_s1068366612020110-Figure7-1.png",
+        "caption": "Fig. 7. Schematic of calibration of friction torque measuring system of SI 03M test machine.",
+        "texts": [
+            " The check involves the following two modes: \u2014when the bending load is directed \u201cUPWARDS\u201d; \u2014when the bending load is directed \u201cDOWNWARDS\u201d. The arrangement of the devices and proving rings is shown in Fig. 6. Before calibration, the bending load control system is calibrated. The load is assigned in the manual control mode. The fine adjustment of the load is carried out using the rod (see Fig. 6). The accuracy of measurement of the bending load is 2%. (7) The error of the friction torque measurement system is determined using the reference weights of the 4th class (GOST 1328\u201382). The arrangement of the device is shown in Fig. 7. The lever is initially equili brated with a weight and then the friction torque meter is halted using the rod. The check is carried out by placing reference weights of the 4th class with masses of 0.1, 0.2, 0.4, 0.6, and 0.8 kg\u2014corresponding to friction torques of 0.5, 1.0, 2.0, 3.0, and 4.0 N m\u2014on the plate. A weight with a mass of 0.1 kg yields a torque of 0.5 N m and that with a mass of 0.2 kg yields a torque of 1.0 N m. The length of the calibration lever arm is L = 510 mm. The accuracy of measuring the friction torque is 2%"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000724_s1064230707040077-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000724_s1064230707040077-Figure4-1.png",
+        "caption": "Fig. 4. The test example.",
+        "texts": [
+            " Hence, the pair of the roots \u03bb = \u00b11 of the system matrix J corresponds to each nonzero eigenvalues of the matrix JG, and the other roots equal zero. It is obvious that in this case the convergence of iterations is not ensured (4.2). In practice, this means that when the stiffness of the force elements grows, the integration step of the motion equations decreases. To illustrate the study we carried out, we consider a simple example of a system of bodies with stiff force interaction. The 2D system consists of three identical pendulums in the form of thin homogeneous rods connected sequentially via kinematic pairs (Fig. 4). The first pendulum is connected with the fixed post. Each pendulum is 1m long and has a mass of 10kg. The gravity force is directed down. The middle of the first pendulum and the end of the last pendulum is connected by a linear bipolar element, whose damping coefficient is denoted by \u03b1. The stiffness of the motion equations increases with the growth of the numerical value \u03b1. We performed the integration by the Park method with the local error tolerance at one step \u03b5 = 10\u20138, and the integration interval was 10 s. The initial position of the system is given in Fig. 4, and the initial velocities are zero. Figure 5 compares the efficiency of calculations when we do not use the Jacobian matrices, use the approximate, and block diagonal Jacobian matrices. We took the number of calculations of the elements of the motion equations Ne as the efficiency index. The increase in the stiffness of equations results in slowing down the integration without the Jacobian matrices. The addition of the approximate Jacobian matrices results in very high efficiency since the length of the integration step is almost independent of the stiffness of equations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003846_12.970309-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003846_12.970309-Figure1-1.png",
+        "caption": "Fig. 1. (a) Model of an azobenzene elastomer. Each network strand consists of N freely-jointed rod-like Kuhn segments of the length l. (b) Two models for network strands with orientation of chromophores parallel (upper panel) and perpendicular (lower panel) to the long axis of a Kuhn segment.",
+        "texts": [
+            " The orientation LC-interactions are taken into account in the framework of the mean-field approximation, whereas interaction of the chromophores with the light is described in the framework of the orientation approach developed in refs. 25-28. We show that in the plane perpendicular to the polarization vector of the light the chromophores can form additionally the LC-order, so that the system becomes biaxial. The biaxial ordering of the chromophores leads to the biaxial deformation of an azobenzene elastomer. An azobenzene elastomer is modeled as an ensemble of polymer chains (network strands) between network junctions, see Figure 1a. Each network strand consists of N freely-jointed rod-like Kuhn segments of the length l bearing azobenzene chromophores. In the present study we consider azobenzene elastomers built from flexible macromolecules, so that each Kuhn segment is assumed to contain only a single azobenzene chromophore. As it was shown by computer simulations, this assumption can be fulfilled for azobenzene polymers built from polyethylene backbones at temperatures higher than the glass-transition temperature.[37] In order to study the influence of chemical structure of network strands on the light-induced deformation of azobenzene elastomers (e.g., influence of the orientation distribution of azobenzene chromophores with respect to the backbones of the network strands) we consider azobenzene elastomers of two limiting structures in which the chromophores are Proc. of SPIE Vol. 8545 854507-2 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 09/23/2015 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx oriented parallel or perpendicular to the long axes of Kuhn segments, see Figure 1b. It is assumed that the chromophores are rigidly attached to the Kuhn segments. These structures can mimic qualitatively the main- and side-chain azobenzene elastomers, respectively. Application of the linearly polarized light to an azobenzene elastomer leads to a complicated quantum-mechanical process of multiple trans-cis-trans photo-isomerization of chromophores. The trans-cis-trans photo-isomerization process is accompanied by alignment of long axes of the chromophores perpendicular to the polarization direction E of the light",
+            "[38] It is a symmetric traceless tensor[38] and, thus, it can be diagonalized in an orthogonal frame of reference denoted here by XYZ: \u239f\u239f \u239f \u23a0 \u239e \u239c\u239c \u239c \u239d \u239b = zz yy xx S S S 00 00 00 S\u0302 , (4) were Sii = [3\u2329cos2\u03b8i\u232a \u2212 1]/2 is the scalar order parameter with respect to the i-axis (i = x, y, z) and \u03b8i is the angle between the long axis of a chromophore and the i-axis. Here the averaging is performed over all chromophores. Note that the orientation of chromophores is symmetric with respect to the transformation E \u2192 (-E) according to the potential given by Equation (1). Due to this symmetry one of the eigenvector of the tensor order parameter should coincide with the vector E. To specify, we choose the x-axis to be directed along the vector E, see Figure 1a. Since \u015c is the traceless tensor (i.e. Sxx + Syy + Szz = 0), only two components are independent. Below we will analyze the light-induced orientation ordering of the chromophores in terms of two scalar order parameters S and \u03bc: S = Sxx and \u03bc = 2(Syy \u2212 Szz)/3. The parameter S is a usual uniaxial order parameter with respect to the polarization direction of the light E, while the parameter \u03bc determines the orientation biaxiality: \u03bc = 0 for the uniaxial orientation of chromophores Proc. of SPIE Vol",
+            " We demonstrate that orientation biaxiality of chromophores results in the biaxial deformation of azobenzene elastomers. Reorientation of chromophores with respect to the polarization direction E results in the reorientation of the Kuhn segments with respect to this vector due to rigid coupling between the chromophores and backbones of network strands. Reorientation of the Kuhn segments leads to network deformation. Thus, under light illumination the network strands change their conformations, and each end-to-end vector b (see Figure 1a) is transformed into a new vector b\u2032. As in a classical theory of rubber elasticity,[41,42] we assume that network strands deform affinely with the bulk deformation of the elastomer because of the constraints of the crosslinks: xxx bb \u03bb=\u2032 , yyy bb \u03bb=\u2032 , and zzz bb \u03bb=\u2032 , (10) where \u03bbx, \u03bby, and \u03bbz are the elongation ratios of the sample along the x-, y-, and z-axes. Due to incompressibility of elastomers, the elongation ratios are related to each other according to the following condition: 1=zyx \u03bb\u03bb\u03bb ",
+            " At the same time, a biaxial ordering, \u2329lx 2\u232a \u2260 \u2329ly 2\u232a \u2260 \u2329lz 2\u232a, results in the biaxial deformation of an elastomer, \u03bbx \u2260 \u03bby \u2260 \u03bbz. The averaged projections of the Kuhn segments on the three principal axes, \u2329l\u03b2 2\u232a, are related to the orientation order of the chromophores attached to them, this relationship being determined by the orientation of chromophores with respect to the main chains. Below, we discuss the light-induced elongations for the main- and side-chain azobenzene elastomers presented schematically in Figure 2b (see upper and lower panels, respectively). In the main-chain azobenzene elastomers (upper panel in Figure 1b) the long axis of a chromophore coincides with the long axis of a Kuhn segment. Thus, the values \u2329l\u03b2 2\u232a are related to order parameters of the chromophores in a simple way: 3 12 2 2 +=\u232a\u2329 S l lx , 6 322 2 2 \u03bc+\u2212= \u232a\u2329 S l ly , and 6 322 2 2 \u03bc\u2212\u2212=\u232a\u2329 S l lz . (17) Using the dependences S(V0/kT, \u03b1) and \u03bc(V0/kT, \u03b1) discussed in Section 3, and using Equations (16) and (17) we have calculated the elongation ratios \u03bbx, \u03bby and \u03bbz as functions of V0/kT for main-chain azobenzene elastomers with very strong orientation interactions, \u03b1 > \u03b1C, when the elastomers are in the LC state at the absence of the light",
+            "[16] Moreover, we have found that the increase of the parameter of the orientation interactions \u03b1 leads to the increase of the magnitudes of both light-induced contraction along the vector E and light-induced elongation along the LC-director of chromophores (i.e. along the y-axis). Furthermore, one can see from Figure 4 that the value of \u03bbz is a monotonically increasing function of V0/kT in the biaxial state. Thus, the main-chain azobenzene elastomers contract along the polarization direction of the light, \u03bbx < 1. In the present subsection we consider the light-induced deformation of side-chain azobenzene elastomers, in which the chromophores are oriented perpendicular to the backbones of the network strands, as presented in the lower panel of the Figure 1b. We assume that the orientation interaction takes place only between the chromophores and the segments of Proc. of SPIE Vol. 8545 854507-7 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 09/23/2015 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx the backbones do not interact with each other. Thus, the Kuhn segments are oriented randomly around the long axes of the chromophores. Averaging over the random orientation of the Kuhn segments around the long axes of the chromophores, we obtain the following relationships between the average projections of the Kuhn segments on the three principal axes \u2329l\u03b2 2\u232a with the order parameters of the chromophores S and \u03bc: 3 1 2 2 S l lx \u2212=\u232a\u2329 , 12 324 2 2 \u03bc\u2212+= \u232a\u2329 S l ly , and 12 324 2 2 \u03bc++=\u232a\u2329 S l lz "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000958_ol.34.003214-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000958_ol.34.003214-Figure1-1.png",
+        "caption": "Fig. 1. (Color online) Structure of bio-inspired fluidic IOL, consisting of a support ring, a lens chamber, and a reservoir. Fluid can be exchanged between these two chambers via eight channels in the support ring. The optical fluid is sealed inside both chambers by two elastic membranes.",
+        "texts": [
+            " In this Letter, we use the bio-inspired fluidic lens for an accommodating IOL, the shape of which can be controlled by ciliary muscle. The results 0146-9592/09/203214-3/$15.00 \u00a9 demonstrate that the fluidic IOL can achieve an accommodation range of 12 D, which is 6 times as high as the state-of-the-art IOLs. The results also suggest that this accommodation range can be achieved with a muscle force and displacement compatible with the physiological conditions of aged eyes, thus promising complete restoration of vision accommodation. The structure of the bio-inspired fluidic IOL is shown schematically in Fig. 1. The spherical periphery defines the equator of the IOL. The device has two membranes, each bonded to the support ring and to each other to form a lens chamber and the equatorial chamber. The membranes and the support ring are made of polydimethylsiloxane (PDMS) because of its desirable optical and mechanical properties and biocompatibility. The ring has a wall thickness of 1 mm and a width of 2 mm. On the side wall of the ring there are eight through holes connecting the 2009 Optical Society of America equatorial reservoir to the lens chamber"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001727_0022-2569(69)90006-8-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001727_0022-2569(69)90006-8-Figure7-1.png",
+        "caption": "Figure 7.",
+        "texts": [],
+        "surrounding_texts": [
+            "268\nOver the practical range - 1 5 ~ < = z 3 < 1 3 5 \u00b0 we conclude that the mechanism is a doublerocker if - 15\" < ~z3 <450 and if 75 \u00b0 <ez3 < 135\u00b0 and a crank-rocker if 45 \u00b0 <:~23 < 75\".\nIf Xz3=45 \u00b0 or 75 ~ the mechanism is a c rank-rocker but suffers from change points. For these values, ~.,3 +:~34 and z - ,3 - :q4 , respectively, are tangential to the curves in Fig. 4, and equation (3.15) reduces to\ncos0 t = + l , cos0 z = ~ l , cos 0 3 = + 1 , cos04=-T-1 A severe constraint has now been applied to the original R - C - C - C mechanism and we are not permit ted a free choice of values for the link lengths a u. Equations (A.I ) - (A.5) reduce to:\n(at z --+ az 3) = (a34 -4-a41 ) $2 sin(cc23 --- 7z34) = ---Sx t sin ~4t $3 = 0 -T-S+ sin(z~z3 -+-~3+) = ___St1 sin ~tz\nThe links a23 and a34 are thus coUinear and parallel to the links a~2 and a+~. If S ~ t = 0 then all links become collinear. For double change points the sine and cosine laws for spatial triangles give the geometry of the Bennett mechanism.\n4. Classification of the R - C - C - R Mechanism R - C - C - R mechanisms have been reported by Dimentberg and Al tman [6]. The eight dual angles for the most general form (Fig. 6) are,\nC212 = 0 +~a12 0 1 = 0 t +8S11 ~23 = 0~--[- 8a23 0 2 - 0 2 + e S 2 c23~,, = 0 + ca34 03 = 03--beS 3 c2+1 = ~ + ea4t 04=0++eS4+ (4.1)\n(i) Displacement equations By substituting (4.1) into equations (2.1)-(2.5) and separat ing into p r imary and dual components we obtain: a34 cos 0+ = (a23 -- a+t ) - a I 2 cos 01 sin 02 =s in 0t $2 sin ~ cos 0z = a3+(cos 01 sin 0+ + c o s ~ sin 0t cos 04) - ( a z 3 - a+l)cos ~ sin 01 + S i t sin = cos 0t sin 03 = sin 04",
+            "269\n$3 sin 7 cos 83 =a l z ( s in 81 cos 8++ cos m cos 8 t sin 84) - (a23 - a 4 t ) cOs ~t sin 83 + $44 sin ~ cos 04 al 2 cos 82 = a34 cos 8+-- (a23 -- a41 ) at z cos 8 t = a34 cos 83 + (a 23 - a+ t) (4.2)\n(ii) Expressions for limit positions We note that the pr imary parts of the general displacement equations (2. I)-(2.5) have reduced to zero or simple relationships such as sin 82 = sin 81 etc. A corresponding spherical mechanism does not exist and therefore we would not expect the pr imary parts of(3. i)-(3.14) to give expressions for limit positions for the overconstrained R - C - C - R mechanism. From the dual parts o f these equations we obtain for an input limit posi t ion:\nCOS 81 - - ( a23 ----- a34 ) - - a+l 1212\ncos 02 = _ (az3 + a 3 4 ) - a41 a12 83=0 ,\n0 + = z , 0 (4.3)\nand f rom the displacement equat ions:\nS 2 = [a 122 - - (a23 ____. a34 -- a+ 1)2]~cot ct -- S i t\n53 = _ [ a 1 2 2 _ (a23 __. a34 _ a 4 t ) 2 ] + c o s e c c( - S++\nand for an output limit posit ion:\n8t = ~ , 0\n82 = O, rc\ncos 83 -- a41 - ( a 2 3 ___a12) 034\ncos 84 = -- a+l - - ( a 2 3 \"}- a12 ) (/34.\nSe = - - [a 342 - - (agt -- ae 3 -T- a l 2)z'] ~ cosec ~(-- S l l\n53 = E a 3,42 - - (a41 - - a23 ~ a 12)2\"] ~ cot c(-- $44 (4.4)\nThe double signs correspond to the two possible input and output limit posit ions illustrated in Figs. 7 and 8.",
+            "270\n(iii) Classification For input limit positions Icos 011(= [cos 021)< 1 and from (4.3) we obtain\na12 ___a41 >a2a ---a34 (4.5)\nSimilarly from output limit positions (4.4) and using Icos 031 ( = t cos 041)< I\na12 q -a4 i < 4 2 3 q-a34 (4.6)\nA crank-rocker mechanism is obtained by satisfying (4.6) since clearly (4.5) is then not satisfied. A rocker-crank is obtained by satisfying (4.5). A drag-link mechanism having change points is obtained by satisfying\na12 +__a4t =a2a -t-a34 (4.7)\n(i.e. (4.5) and (4.6) are not satisfied).\nEquation (4.7) gives\na12=a:3 a34=a41 (4.8)\nSpecial case. A well known type of R -C-C-R mechanism is obtained by substitution of\na23=a41 =0 in (4.1)\nInput limit inequalities (4.5) reduce to\na12 ) --~-434 (4.9)\nOutput limit inequalities (4.6) reduce to\na12 ~ ___a34 O.lO)"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000689_09544100jaero206-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000689_09544100jaero206-Figure3-1.png",
+        "caption": "Fig. 3 Coanda effect theory",
+        "texts": [
+            " Aerospace Engineering at YORK UNIV LIBRARIES on July 12, 2012pig.sagepub.comDownloaded from Co-flow fluidic thrust vectoring system relies on a phenomenon known as Coanda effect, which falls back on viscosity. It states that fluid and gaseous jets have a natural tendency to attach to the wall, which is projected close to them and follow the convex curvature of the solid boundary [34]. It happens due to reduction of surface pressure owing to a vortex action as the liquid or gas passes over the boundary. The co-flowing mechanism is schematically shown in Fig. 3. The pressure at the bond surface is less than the pressure at the primary jet. For a particle traveling along the curved surface, the centrifugal force must be equivalent in magnitude to the pressure forces binding the fluid to the surface. This explanation can be mathematically simplified as Centrifugal force = m V 2 R ; m = \u03c1(R \u03b8)Lt (23) Suction force = P(R\u03b8)L; P = PT \u2212 Ps (24) by equality of these two forces P = \u03c1V 2t R (25) It means that the pressure difference depends on the jet sheet thickness and Coanda surface radius"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003641_s00775-013-1011-7-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003641_s00775-013-1011-7-Figure5-1.png",
+        "caption": "Fig. 5 Cyclic voltammetry of a freshly prepared TvNiR film in 0.1 mM nitrite, pH 7.0. a continuous lines show the initial cyclic voltammogram (bold line) and four subsequent cyclic voltammograms from a TvNiR film. Broken lines show the voltammograms from an equivalent experiment in the absence of TvNiR. Cyclic voltammetry was performed at 30 mV s-1 with electrode rotation at 3,000 rpm in 20 mM Hepes, 100 mM NaCl, 293 K. b the catalytic current\u2013potential profiles of TvNiR obtained from the data in a as the difference between the currents recorded in the presence and the absence of TvNiR. Bold lines are for the first cyclic voltammogram, and arrows indicate the scan direction. Crosshairs indicate the zero of current and potential. S.H.E. standard hydrogen electrode",
+        "texts": [
+            " Substrate mass transport to enzyme molecules adsorbed deep within the mesoporous structure limits the catalytic current and masks the intrinsic properties of the enzyme [35]. Instead both enzymes were adsorbed on PGE electrodes that were rotated rapidly to further facilitate substrate delivery to the electrode surface. Cyclic voltammetry of freshly polished PGE electrodes exposed to a few microlitres of TvNiR, placed in 0.1 mM nitrite, and rotated at 3,000 rpm gave rise to negatively signed, and so reductive, catalytic currents below -100 mV (Fig. 5a). The catalytic response did not change on addition of 100 mM phosphate but it was lost on introducing 500 lM cyanide (not shown) (cyanide is toxic and extreme care should be taken in its handling and disposal). These observations correlate with the catalytic properties of TvNiR resolved by spectrophotometric assays using dithionite-reduced methyl viologen as an electron donor [22]. It was concluded that TvNiR, like EcNrfA [20, 36], adsorbs on PGE electrodes in an electrocatalytically active form that retains the characteristic properties of the enzyme in solution",
+            " This includes the possibility of a distribution of orientations on the surface, many of which are incapable of direct electron exchange with the electrode, and it prevents the calculation of turnover numbers for nitrite reduction. Nevertheless, the variation in catalytic current with electrochemical potential, time, and pH provides much insight into the catalytic performance of the enzymes as described below. The catalytic properties of TvNiR are more apparent when the voltammetric response of the bare electrode is subtracted from that displayed by the enzyme film under otherwise identical conditions (Fig. 5b). The catalytic current magnitudes of the first reductive scan from ?200 to -600 mV (Fig. 5b, bold line) are smaller than those resolved on returning to more positive potentials (Fig. 5b, thin line). The catalytic current\u2013potential profiles of the subsequent cyclic voltammograms do not show this hysteresis (Fig. 5b, broken lines). The profiles for cyclic voltammograms 2\u20135 are independent of the scan direction aside from a first-order loss of catalytic current magnitude over time that is typical of such experiments and is attributed to loss of electrocatalytically active enzyme. Significantly, the catalytic current\u2013potential profiles of cyclic voltammograms 2\u20135 resemble that seen on returning to positive potentials during the first voltammogram. Taken together, the cyclic voltammetry data describe an irreversible reductive activation, on the timescale of the PFE experiment, of the TvNiR film on scanning from ",
+            " This has the advantage of avoiding the intimate linkage of time and electrochemical potential that is intrinsic to cyclic voltammetry. The electrode potential was stepped between defined values, and the subsequent relaxation of the catalytic current reported on the rate of any reductive activation that occurred. A typical result is illustrated in Fig. 6a. When the experiment is initiated at ?140 mV with the electrode rotating rapidly in 0.1 mM nitrite, there is little flow of current, as expected from cyclic voltammetry (Fig. 5). The small drop of current noted in the first 20 s arises from charging of the electrical double layer. When the potential is stepped to -160 mV, there is an immediate and significant increase in the catalytic current, as anticipated from cyclic voltammetry (Fig. 5). The current due to double-layer charging decays over 10 s, after which time a constant catalytic current is measured. Stepping the potential to -270 mV elicits an immediate increase in catalytic current, again as expected from the cyclic voltammetry, followed by a further significant increase in the catalytic current over 120 s that describes the reductive activation of the protein film. Returning the electrode potential to -160 mV showed the catalytic current had increased by approximately 30 % from that prior to reductive activation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure6-1.png",
+        "caption": "Fig. 6. Torsion loadcase: forced response",
+        "texts": [
+            " This is a general approach which could be used for several load cases. The forced response can be done over a certain frequency range (e.g.: 5-30 Hz) and the static stiffness can be extracted by fitting/extrapolating the curve towards 0 Hz. This fitting will be necessary since at the very low frequencies, the results will be very noisy when working with measured FRF data. Inverse force identification Although the approach of a forced response should be suitable for most load cases, it is not appropriate to calculate the static torsional stiffness. As shown in Fig. 6 two opposite forces are applied at the front domes while the rear domes are clamped. Since the vehicle body is only constrained at the rear domes, it still has a rotational rigid body mode. Using Virtual.Lab [5] as simulation platform a forced response calculation has been performed, calculating the displacement of the front domes from 0 to 3 Hz. This calculation leads to very high displacements around 0 Hz which makes it impossible to have an accurate estimation of the static torsional stiffness (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003697_acc.2013.6579809-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003697_acc.2013.6579809-Figure1-1.png",
+        "caption": "Fig. 1. Basic Engagement Geometry",
+        "texts": [
+            " Keeping in mind that 3-D engagement scenario is more practical and the PN guidance law is the most widely used one in practice, this paper analyzes a generic PN guidance law, that uses standard PPN (Pure proportional navigation) and RetroPN guidance laws in a 3-D engagement scenario based on the initial engagement geometry and terminal engagement requirements, against higher speed nonmaneuvering targets to control terminal impact angle. Results are obtained on the set of achievable impact angles and the conditions on the navigation constant to achieve them based on initial engagement geometry in the capture zone of this proposed generic PN guidance law. Consider the interceptor-target engagement geometry shown in Fig. 1. The LOS fixed reference frame consists of er, et ( e\u0307r/\u03c9 ) , e\u03a9 ( \u03a9/\u03c9 ) as the three orthogonal unit vectors [19], [21], [22]; where, er is the unit vector along the LOS and \u03a9, the angular velocity of the LOS, is orthogonal to the LOS and is given by, \u03a9 = \u03c1\u22121er \u00d7 (vT \u2212vM) (1) Here, \u03c1 is the range distance and vM, vT are the interceptor and target\u2019s velocity vector, respectively. The magnitude of the angular velocity of LOS vector with sign is defined as, \u03c9 = { \u2016\u03a9\u2016 ,when er moves anticlockwise \u2212\u2016\u03a9\u2016 ,when er moves clockwise (2) The dynamics of the engagement is expressed in terms of the differential equation of the relative position vector r (LOS vector) as given below"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure2-1.png",
+        "caption": "FIGURE 2. The screw used with a toothed wheel (Mechanics 2.6). On the left is Drachmann\u2019s drawing (The Mechanical Technology, p. 61) from Ms B; on the right is the figure from Heronis Alexandrini opera, vol. II, p. 110.",
+        "texts": [
+            " these devices as belonging to a special class, and was quite independent of any theoretical understanding of their operation. Heron begins his account of the five powers with a description of their construction and use (2.1\u20136) that reveals a close familiarity with practitioners\u2019 knowledge.15 The construction of the wheel and axle (2.1), the compound pulley (2.4), and the screw (2.5\u20136) is described in detail. Two uses of the screw are described: with a wooden beam or \u0301 o (Fig. 1), and with a toothed wheel (Fig. 2).16 The account employs a good deal of specialized terminology for the mechanical powers and their parts; a number of these technical terms are explicitly flagged as such using the Greek word \u0302 , \u201cto be called\u201d.17 A striking feature of Heron\u2019s account is the statement of rough, non-quantitative correlations describing the behaviour of the five powers. Thus instead of a precise formulation of the law of the lever as a proportionality between forces and weights, Heron remarks that \u201cthe nearer the fulcrum is to the load, the more easily the weight is moved, as will be explained in the following\u201d (2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001026_j.ijadhadh.2008.08.001-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001026_j.ijadhadh.2008.08.001-Figure3-1.png",
+        "caption": "Fig. 3. Contact between a fiber and a rough surface: (a) the deformation of the",
+        "texts": [
+            " The effect of tip enlargement will be addressed later. fiber tip due to surface asperity amplitude (d) produces an elastic force that tends to detach the fiber from the surface. (b) Schematic of the fiber before deformation when elastic force is applied to the fiber. In the previous section, it was assumed that fibers come into contact with an ideal smooth surface. This type of contact will not cause elastic deformation in fibers. When coming into contact with a rough surface, the fibers deflect elastically to adjust themselves as shown in Fig. 3. To discern the relation between forces applied to the fiber tip and fiber deformation, a more general theory is used [35]. Consider that fibers are protruding out of the backing at angle y (0oyo90). If d represents surface roughness height at the point of contact, a normal force (F) and a shear force (V) are applied to the fiber tip. The magnitude of shear force V is limited by Coulomb friction [36] and is equal to where m is the friction coefficient. V and F result in a compressive force along the fiber and a bending component normal to the fiber, obtained from the following equations: FCmp \u00bc F sin y\u00fe mF cos y (9"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003056_ssd.2011.5767475-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003056_ssd.2011.5767475-Figure3-1.png",
+        "caption": "Figure 3. Multi-Layer neural network based wind speed estimation algorithm",
+        "texts": [
+            " The observer model is represented by: ( )u u mecu em mec mec uu d X A X BT G dt C X \u2227 \u2227 \u2227 \u2227 \u2227 = + + \u03a9 \u2212 \u03a9 \u03a9 = (22) where mec u w X T \u03a9 = ; 10 0 0 uA J = ; 1 0 T B J = \u2212 ; [ ]1 0uC = and G is the matrix gain observer which can written as: 1 2 g G g = . To ensure the observer stability we determine the matrix G coefficients so that )( uu GCA \u2212 is stable. In equation (22) Tem and mec, which are not measurable in the control algorithm, are substituted by the estimated values obtained from the neuronal observer. In the proposed algorithm the wind speed is estimated by using the estimated aerodynamic torque and rotor speed. We use three layers neural network. As shown in figure 3 the network proposed is constituted of 2 inputs corresponding to the estimated aerodynamic torque and rotor speed and a single output corresponding to the estimated wind speed. The latter is used after to impose the optimal output power reference, by using an MPPT block, based on knowledge of the WECS dynamics and power characteristics. Figure 4 gives the global sensorless control for wind generation system based on doubly fed twin stator induction generator using Mutli Layer Neuronal Network. igure 4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002299_978-3-642-22700-4_42-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002299_978-3-642-22700-4_42-Figure1-1.png",
+        "caption": "Fig. 1 a DH representation of legs for the Nao robot kinematic simulation; b mass and geometrical model for the Nao robot",
+        "texts": [
+            " Closed-form solution of the non-linear equations of the robot manipulators with limited degrees of freedom is introduced as an alternative method of solving inverse kinematics equations. This method is also employed to carry out the solution of MRL team Nao robot\u2019s kinematic equations with utilization of Pieper\u2019s solution [23] for developing transformation and rotation matrices. This method and its modification is described in this section as an approach for solving related equations which enables the approach to compute the same hip Yaw-Pitch joint variable for both legs. DH frame assignment and related coordinate sequences are shown in Fig. 1, and associated parameters are mentioned in Table 1. DH parameters for the right leg is assumed the same as the left one except a6 which is p=4. Homogeneous transformation 0 6T which describes position and orientation of frame {6} relative to frame {0} is introduced as Eq. 8 0 6T \u00bc 0 6R 0p6ORG 0 1 \u00f08\u00de Position vector 0p6ORG expresses coordination of frame {6} origin relative to frame {0} in the homogeneous transformation matrix. Joint angles h1; h2; h3 could also be computed by using 0p6ORG; then the last three joint angles of each leg, h4; h5; h6 are approximated using 0 6R which is rotation matrix explaining orientation of frame {6} relative to {0}",
+            " 13 Ti \u00bc Ri;3 3 Li;3 1 O1 3 1 \u00f013\u00de Ri s are the basic rotation matrix for all 12 joints about three axes and defined in Eq. 14. In addition, Li s are joints\u2019 vectors as shown in Table 3, Appendix. Basic rotation matrix for three-dimensional transformation in xyz coordinates are as follows: Rx;h \u00bc 1 0 0 0 cos h sin h 0 sin h cos h 2 6 4 3 7 5; Ry;h \u00bc cos h 0 sin h 0 1 0 sin h 0 cos h 2 6 4 3 7 5; Rz;h \u00bc cos h sin h 0 sin h cos h 0 0 0 1 2 6 4 3 7 5 \u00f014\u00de Table 2 represents all 12 joint numbers and their axis of rotation are as depicted in Fig. 1a. Global position of links\u2019 center of mass vectors are defined from the successive multiplication of transformation matrix in Eq. 13 and the local position vectors tabulated in Table 3, Appendix I as: Pmj \u00bc Yn i\u00bc1 Ti ! Lmj \u00f015\u00de Pmi \u00bc xmi ymi zmi 1h iT \u00f016\u00de where i starts from 1 and ends to the number of joint n before mass i. A dynamic model is developed to determine the joints\u2019 angular acceleration for which the mass and diagonal elements of the inertia tensors of each Nao\u2019s link are listed in Table 4 and schematically illustrated in Fig. 1b. A movement study of the MRL team Nao robot and ZMP estimation will be done after determination of joint values. The presented dynamic model is for the SSP and will be used as a fundamental calculation for the DSP which is the authors\u2019 next research subjects. SSP dynamic equations are derived by the assumption of the Lagrange-d\u2019Alembert Eq. 17 which relates the external force Fi on the Nao links for the single support phase and variations of Lagrangian L relative to the system state ri. Name Axis Joint d dt oL o_rsn oL orsn \u00bc Fn \u00f017\u00de The state vector rsf g consists of the position vectors and the angles as: rsf g \u00bc x1 y1 z1 x5 y5 z5 h1 h12h iT \u00f018\u00de The Lagrangian L is defined as the difference between a system\u2019s kinetic energy and potential energy and could be written as follows: L \u00bc XNumber of Links n\u00bc1 1 2 mn\u00f0 _x2 n \u00fe _y2 n \u00fe _z2 n 2gzn\u00de \u00fe 1 2 ~xT n In~xn \u00f019\u00de In which mn is the mass of nth link, In is the inertia tensor of nth link around base frame, and ~xn is the angular velocity vector of the link n around base frame",
+            " Assumption of left leg as the non-supporting leg justifies small values of calculated T9 and T10 values in comparison with the corresponding joints on the right leg and zero values for T11 and T12. Phase estimator calculates the phase parameters for SSP and DSP states, and then joints\u2019 reference angles are used in the forward kinematic model together with the phase estimator to produce all links\u2019 state position, velocity and acceleration vectors rsf g; _rsf g; \u20acrsf g. Generalized position vectors of links and five masses shown in Fig. 1b are estimated by Eq. 15, then velocity and acceleration values are achievable with numerical differentiation. Figures 8 and 9 demonstrate acceleration of two assumed mass on the torso and the right leg shown as m2 and m3 in Fig. 1b. State position vectors and actuator torques are executed in the dynamic model to generate real joints\u2019 angles and velocities. The reference joint angles are compared with the estimated ones to build up a reliable controller such as adaptive PID or fuzzy logic controller. Torque control is this group\u2019s next research topic in order to reach a smooth Nao bipedal motion on the slopes. All plotted curves are based on single support assumption of the right leg, but the transient mode and successive leg changes have not been considered in this simulation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001103_insi.2009.51.8.426-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001103_insi.2009.51.8.426-Figure3-1.png",
+        "caption": "Figure 3. Defect on (a) outer race, (b) inner race and (c) ball",
+        "texts": [
+            " It should be noted that the outer race defects were not simulated but occurred naturally due to the excessive load applied (larger than the dynamic load rating), the dimensions of which are detailed in Table 1. Table 2 presents defect dimensions on the inner race which were artificially seeded. During the tests, inspection of the inner and outer rings for measurement of the size of defect area was performed. The test was interrupted every 400,000 cycles to measure the defect size. Some sample defects in the bearing elements are shown in Figure 3. Table 1. Estimation of defect length on outer race through LCR values Running Cycle (million cycle) LCR (No/sec) Estimated Defect Length (mm) r.m.s. (m/sec2) Actual Defect Length (mm) 1.1 8129 0.7 21 1 1.3 9026 8.8 32.8 9.25 1.4 9138 9.9 39.7 10.7 1.5 9189 10.2 40.4 11.1 1.6 9268 11 43.2 12.25 Each bearing element has a characteristic rotational frequency. With a defect on a particular bearing element, an increase in vibration energy at this element\u2019s rotational frequency may occur. These characteristic defect frequencies can be calculated from kinematic considerations, ie the geometry of the bearing and its rotational speed[15]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003734_j.engfailanal.2012.11.013-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003734_j.engfailanal.2012.11.013-Figure2-1.png",
+        "caption": "Fig. 2. General dimension of the spur gear. (a) Positioning of the holes for handling the gear, (b) location of the holes adjacent to fracture surface, (c) alignment between matching holes at the fracture.",
+        "texts": [
+            " In order to access the conformity of the gear to the design specification chemical analysis, hardness and microstructural evaluation were performed. The chemical composition was evaluated at the core of the gear using an optical spectrometer Spectra \u2013 Spectrolab. Vickers hardness profile was taken with 500 g load with 0.05 mm increments from the surface of the tooth to the core. Optical microscopy was used to evaluate the case and core microstructure. The fracture surface of the gear was examined by both naked eye and stereo microscope in order to determine the fracture initiation point and the general fracture surface characteristics. Fig. 2 shows a basic drawing of the fractured gear. The design project included several holes for handling and transport of the large gear. Those holes where placed 90 apart and a perfect alignment between matching holes were not achieved as can be seen in Fig. 2c. Stress distribution within the gear was determined by the finite elements method. A solid with dimension of the gear was generated including two of the machined holes used for the gear handling. The intersection between the two holes resulted in an elliptical flaw with a larger diagonal of approximately 6 mm. For the modeling the gear was restricted by its inner diameter. The load applied on modeling was derived from the gear operating torque. It was distributed over three teeth adjacent to the flaw and applied in the pressure line direction on a surface of 3670 mm2 over each the three teeth, so that the tooth just above the flaw experienced a stress of 80 MPa, and its neighboring teeth experienced stresses of 65 MPa",
+            " The numerical simulations based on the geometry of the defect and service loads applied to the gear indicated that the stress intensity at the defect lies in the stable fatigue crack propagation region. To prevent future failures, the adoption of an inspection plan after manufacturing that also consider the region of the holes is recommend. Furthermore, the position of the holes for handling the gear is too close to a region under the gear teeth where stresses are concentrated during gear operation. The possibility of slightly decreasing the 689 mm distance between hole centers (see Fig. 2) so that they move away from the region of higher stress should be evaluated. A reduction to 600 mm, for example, should not significantly alter the gear transportation logistics but would have a decisive influence in terms of the stress level at the singularity. The authors would like to thank the financial support of CAPES, CNPq and FINEP. [1] Moorthy V, Shaw BA. Contact fatigue performance of helical gears with surface coatings. Wear 2012;276\u2013277:130\u201340. [2] Glodez S, S raml M, Kramberger J. A computational model for determination of service life of gears"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000485_j.fss.2008.03.021-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000485_j.fss.2008.03.021-Figure13-1.png",
+        "caption": "Fig. 13. Two-link flexible-joint robot tracing a square.",
+        "texts": [
+            " The lowest natural frequency of the robot is about 4 Hz without the payload. This reduces to approximately 1 Hz with the addition of the payload. The control software runs at 2000 Hz, including the state-feedback control that augments the CMAC. The CMAC provides outputs, and updates its weights, at a frequency of 20 Hz. The experiments investigate the performance trade-off of traditional robust weight update methods and the new proposed method using CMAC direct adaptive backstepping control. Denoting the two link angles by \u2208 R2 (Fig. 13) and the rotor angles by \u2208 R2, the general form of the dynamics become B( 2)\u0308 = F( , \u0307) + K( \u2212 ), (45) J\u0308 = \u2212D(\u0307) + K( \u2212 ) + u, (46) where B and J represent inertias, K contains spring constants, D is rotor friction, and F includes centripetal forces, Coriolis forces, gravity, and link friction. The states in the state-space representation are x1 = 1, x2 = \u0307, x3 = , x4 = \u0307. (47) Adaptive backstepping control allows the tip to follow a desired trajectory (in this case a square), given in terms of desired link angles d and velocities \u0307d"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001199_0278364907085565-Figure19-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001199_0278364907085565-Figure19-1.png",
+        "caption": "Fig. 19. Atomic modules of P2MR: (a) a module with a single degree of freedom developed during the project (b) a commercial wrist with two degrees of freedom acquired from Shunk, DE (c) a pair of modules with a single degree of freedom from the first generation of P2MR.",
+        "texts": [
+            " Indeed, any modular robot realized for supporting self-reconfiguration (from the electromechanical point of view), by adopting the proposed selforganizing technique, is in a condition to easily operate any module re-arrangement for autonomously varying its shape. A recently concluded research project led to the development of the second generation of P2MR (Plug & Play Modular Robot)10 , a prototype of a modular robot exhibiting the previously described self-coordinating behavior. In its current version it is made up of four heterogeneous modules (see Figure 19), which can be easily assembled together in several alternative ways (see Extension 6) to obtain a class of manipulators with up to five degrees of freedom characterized by different workspaces (see Figure 20). Although the composing modules present highly diversified mechanical characteristics and mount different servos, they share a common element: all of them are equipped with the same PCU, based on a 24 MHz 16-bit microcontroller of the C167 family from Infineon (see Figure 21). Every PCU is interfaced with the corresponding low-level joint controller via an RS-232 serial data link and can communicate with the 10"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003143_1.4000941-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003143_1.4000941-Figure1-1.png",
+        "caption": "Fig. 1 Schematic of a basic journal",
+        "texts": [
+            " Chang epartment of Mechanical and Nuclear Engineering, ennsylvania State University, niversity Park, PA 16802 his paper presents an analytical model for the basic design calulations of plain journal bearings. The model yields reasonable ccuracy as compared with published numerical solutions under he same conditions. The principles and procedures of the formuations are presented along with accuracy analyses. DOI: 10.1115/1.4000941 eywords: journal bearing, design calculations, analytical model Formulation of the Model Figure 1 shows a schematic of a basic journal bearing in a teady-state configuration. The lubricant is supplied from the top egion of the bearing, referred to as the inlet. The hydrodynamic ction generates pressure in the lubricant, primarily in the converent part of the journal-bearing gap, to counteract the load thereby eparating the journal surface from the bearing surface with a thin ubricant film. The hydrodynamic pressure eventually terminates n the divergent part of the gap, referred to as the outlet, and the ubricant film ruptures into streamers"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001838_cca.2010.5611256-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001838_cca.2010.5611256-Figure1-1.png",
+        "caption": "Fig. 1. The testbed for modeling research, 3DOF are locked.",
+        "texts": [
+            " Jun Wu is with the School of Information Science & Engineering, Central South University, Changsha, Hunan 410083, China wujun@csu.ac.cn Hui Peng is with the School of Information Science & Engineering, Central South University, Changsha, Hunan 410083, China huipeng@mail.csu.edu.cn Qing Chen is with the School of Information Science & Engineering, Central South University, Changsha, Hunan 410083, China coolchen302@qq.com is a big challenge itself. In addition, quadrotors may have different configurations, and some of them may be difficult to be taken apart (See Fig. 1). Third, in different flying postures, the coupling dynamics among the outputs are also varying, uncertain nonlinear terms like aerodynamic friction and blade flapping can hardly be taken into account. Therefore, the physical models are usually simplified, sometimes may be rough and inaccurate. To overcome the disadvantages and restricts of the physical models, a new method is necessary. In order to cope with the influence of physical modeling issue, a lot of researches have been done, and some methods have been proposed",
+            " To get an optimal control law and lower online computational burden, a state-feedback gain-scheduling LQR controller based on the identified MIMO RBF-ARX model is designed, and then the performances and contrasts with physical model are also illustrated. The results given in the paper show the advantages of the novel method. In this section, we shall first briefly introduce the quadrotor\u2019s configuration and its physical model-based LQR controller design. The quadrotor to be considered, which is shown in Fig. 1, has 4 propellers, 3 of which are horizontally mounted to control its pitch and roll rotations while the last one is vertically mounted to control its yaw rotation. This quadrotor has 3 outputs and 4 inputs. The outputs are the pitch angle, roll angle, and yaw angle, while the inputs are 978-1-4244-5363-4/10/$26.00 \u00a92010 IEEE 1731 the control voltages of the 4 propellers\u2019 motors equipped at the 4 ends of the quadrotor. The coordinate of the quadrotor is shown in Fig. 2, where Fx(x = f, l, r, b) denotes the thrust forces generated by 4 propellers, its suffixes mean its locations which are front, left, right, and back"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000711_have.2008.4685294-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000711_have.2008.4685294-Figure1-1.png",
+        "caption": "Fig. 1. (a) Spherical burr bone tool, (b) Tool geometry and cutting forces on spherical burr bone tool.",
+        "texts": [
+            " The bone volume is extracted directly from the CT data as a contiguous array of voxels. Our volumetric model can perceive the interior structure and the mechanical properties of heterogeneous bone. Hardware accelerated 3D texture-based volume rendering is used to display the bone and to visually remove the bone material due to drilling simulation in real-time. Stability of the contact is achieved through 978-1-4244-2669-0/08/$25.00 \u00a92008 IEEE a sample-estimate-hold approach [11] in order to remove the excess energy in the tool-bone interaction. Fig. 1(a) shows a spherical tool utilized in almost all bone cutting applications. The geometry and the cutting forces on the spherical milling cutter are shown in Fig. 1 (b). The force on each cutting tool element has three components (dFt, dFr and dFa). dFt is the main cutting force due to the bone orthogonal cutting. dFr is a frictional force on the rake surface. dFa is another frictional force component due to the effect of helix angle of cutting edges [10]. Other parameters of Fig. 1 (b) will be explained shortly. The total force expressed in the Cartesian coordinate system is then obtained by multiplying by the spherical transformation matrix T: \u23a1 \u23a3 Fx(j) Fy(j) Fz(j) \u23a4 \u23a6 = \u23a7\u23a8 \u23a9 N\u03c6\u2211 i=1 Nz\u2211 k=1 T (i, j, k) \u23a1 \u23a3 Kt Kr Ka \u23a4 \u23a6 t(i, j, k)dz \u23ab\u23ac \u23ad (1) where N\u03c6 is the number of cutting edges and Nz is the number of the discretized elements of a cutting edge. The height of each element is dz as shown in Fig. 1(b). Assuming N\u03b8 time intervals as each cutting edge travels the angular spacing of \u03c6, we would like to calculate the force at the jth fraction of \u03c6. The higher the N\u03b8, the more accurate our results will be. t(i, j, k) is the chip thickness in a cutting edge i at the jth fraction of \u03c6. Chip thickness is the most dominant factor in computation of the cutting forces and represented voxel-based in the next section. However, the chip thickness can be derived analytically and is affected by the feed rate of the tool [10]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002850_0022-4898(65)90022-4-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002850_0022-4898(65)90022-4-Figure12-1.png",
+        "caption": "FIG. 12. Schematic drawing of tracti(m measuring device.",
+        "texts": [
+            " It should be noted that the external force S (Table 3) may act as a driving force (P > S) or a towing force ( S > P; S(r/R)=P) or both a driving and a towing force ( S ~ P ; S(r/R)--P). In the case of S=P, the total driving force is using to overcome resistance of self-propelled motion, In the tests, the force 2/' was measured directly by an electrical strain gauge which was placed on a lever of the cable-rollers. The force came from changes of load (that were done in turn). The force 2S was measured by means of the strain gauge which was placed on a lever of cable-drum of the winder. The whole device, shown in the schematic drawing (Fig. 12) was placed on a soil bin. The bin was filled with dry and loose sandy loam with a moisture content of 5 per cent. The soil was cultivated and press rolled before each measurement. Rotation of the 4 0 A . S O L T Y N S K I E N N 0 0.2 0\"4 0-6 0.8 I'0 I-2 P, k g / c m z 0 0.2 0.4 0.6 0,8 i p, kg / cm 2 ]'0 I-2 J.4 I-6 b8 0 0\"~ 0\"4 0\"6 0-8 \", _-.\\ [st. series of m e a s u r e m e n t s 0 ~ ~ . . . . . . . . . . . . 0\"2 ~ o 0 . X % % 0-6 0 ' 8 J.O P4 2 nd. series of m e a s u r e m e n t s b- 6 ' # I t "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002764_0954406211413520-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002764_0954406211413520-Figure1-1.png",
+        "caption": "Fig. 1 A cross section of a floating ring bearing",
+        "texts": [
+            " Numerical simulation is used in section 6 for a selected set of bearing design parameters to confirm the results of the analytical investigation. To derive the equations of motion for the journal and floating ring, Newton second law of dynamics is applied. A symmetric rigid rotor supported by two identical floating ring hydrodynamic bearings is considered. The rotor is assumed to be perfectly balanced and both bearings have the same motions so that only one bearing is accounted for in the analysis. Figure 1 shows a cross section of a floating ring bearing. The journal of radius R1 rotates about its centre O1 at a constant angular velocity !1 inside a floating ring of internal radius R2 and an external radius R3. The fixed bush has an internal radius R4. In a hydrodynamic regime, an internal oil film is formed between the journal and the floating ring and an external oil film is formed between the floating ring and the bearing bush. Due to the difference between the friction torques applied by the internal Proc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000701_13506501jet399-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000701_13506501jet399-Figure1-1.png",
+        "caption": "Fig. 1 Squeeze film configuration of parallel rotating discs",
+        "texts": [
+            " Applying the Stokes couple-stress motion equations together with the continuity equation, the non-Newtonian Reynolds-type equation including the rotational inertial effects is derived and applied to predict the squeeze film performance. Comparing with the conventional Newtonian non-rotational flow, the combined effects of non-Newtonian couple stresses and rotational inertia upon the squeezing film characteristics are presented for different values of the nonNewtonian couple stress parameter, the rotational parameter, and the centrifugal force-load ratio. Figure 1 shows the squeeze film configuration of two parallel circular discs with radius a. The upper rotating disc of angular speed u is approaching the lower fixed one with a squeezing velocity (\u2212dh/dt). The lubricant between the discs is taken to be an incompressible Stokes couple stress fluid. Assume that the thin-film theory of lubrication by Williams [1] is applicable, and the body forces and body couples are absent. Then the continuity equation and the equations of motion considering the rotational inertia for an incompressible couple stress fluid from Stokes [12] are given by 1 r \u2202(ru) \u2202r + \u2202w \u2202z = 0 (1) \u2212\u03c1 v2 r = \u2212\u2202p \u2202r + \u03bc \u22022u \u2202z2 \u2212 \u03b7 \u22024u \u2202z4 (2) 0 = \u03bc \u22022v \u2202z2 \u2212 \u03b7 \u22024v \u2202z4 (3) \u2202p \u2202z = 0 (4) where \u03bc is the shear viscosity, and \u03b7 is a new material constant with the dimension of momentum"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002522_acc.2010.5531598-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002522_acc.2010.5531598-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of nonholonomic autonomous vehicle.",
+        "texts": [
+            " Using the above framework, multi-vehicle missions can be handled in either a centralized or decentralized fashion. For centralized planning, the vehicle dynamic and constraint equations are expanded to include all vehicles. In the decentralized approach, an optimal control problem is solved for each vehicle, taking into account the available information about the others. In the following sections we illustrate how a wide variety of single and multi-vehicle trajectory planning scenarios can be solved using the above formulation. B. Vehicle Model A schematic of the UGV for mathematical modeling is shown in Fig. 1. The UGV has a front wheel steering arrangement, which introduces nonholonomic constraints to the system. The state equations describing the motion of the UGV are as follows: x\u0307 = d dt  x y \u03b8 v \u03b3  =  v cos(\u03b8) v sin(\u03b8) v L tan(\u03b3) a \u03c9  (2) where variables, x, y, and \u03b8 give the position and orientation of the vehicle, measured with respect to the center of the rear axle. The vehicle speed is given by v, L is the vehicle wheelbase and \u03b3 is the steering angle. The control variables are u = [a, \u03c9]T , where a is the vehicle acceleration and \u03c9 is the steering rate"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001727_0022-2569(69)90006-8-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001727_0022-2569(69)90006-8-Figure8-1.png",
+        "caption": "Figure 8.",
+        "texts": [],
+        "surrounding_texts": [
+            "269\n$3 sin 7 cos 83 =a l z ( s in 81 cos 8++ cos m cos 8 t sin 84) - (a23 - a 4 t ) cOs ~t sin 83 + $44 sin ~ cos 04 al 2 cos 82 = a34 cos 8+-- (a23 -- a41 ) at z cos 8 t = a34 cos 83 + (a 23 - a+ t) (4.2)\n(ii) Expressions for limit positions We note that the pr imary parts of the general displacement equations (2. I)-(2.5) have reduced to zero or simple relationships such as sin 82 = sin 81 etc. A corresponding spherical mechanism does not exist and therefore we would not expect the pr imary parts of(3. i)-(3.14) to give expressions for limit positions for the overconstrained R - C - C - R mechanism. From the dual parts o f these equations we obtain for an input limit posi t ion:\nCOS 81 - - ( a23 ----- a34 ) - - a+l 1212\ncos 02 = _ (az3 + a 3 4 ) - a41 a12 83=0 ,\n0 + = z , 0 (4.3)\nand f rom the displacement equat ions:\nS 2 = [a 122 - - (a23 ____. a34 -- a+ 1)2]~cot ct -- S i t\n53 = _ [ a 1 2 2 _ (a23 __. a34 _ a 4 t ) 2 ] + c o s e c c( - S++\nand for an output limit posit ion:\n8t = ~ , 0\n82 = O, rc\ncos 83 -- a41 - ( a 2 3 ___a12) 034\ncos 84 = -- a+l - - ( a 2 3 \"}- a12 ) (/34.\nSe = - - [a 342 - - (agt -- ae 3 -T- a l 2)z'] ~ cosec ~(-- S l l\n53 = E a 3,42 - - (a41 - - a23 ~ a 12)2\"] ~ cot c(-- $44 (4.4)\nThe double signs correspond to the two possible input and output limit posit ions illustrated in Figs. 7 and 8.",
+            "270\n(iii) Classification For input limit positions Icos 011(= [cos 021)< 1 and from (4.3) we obtain\na12 ___a41 >a2a ---a34 (4.5)\nSimilarly from output limit positions (4.4) and using Icos 031 ( = t cos 041)< I\na12 q -a4 i < 4 2 3 q-a34 (4.6)\nA crank-rocker mechanism is obtained by satisfying (4.6) since clearly (4.5) is then not satisfied. A rocker-crank is obtained by satisfying (4.5). A drag-link mechanism having change points is obtained by satisfying\na12 +__a4t =a2a -t-a34 (4.7)\n(i.e. (4.5) and (4.6) are not satisfied).\nEquation (4.7) gives\na12=a:3 a34=a41 (4.8)\nSpecial case. A well known type of R -C-C-R mechanism is obtained by substitution of\na23=a41 =0 in (4.1)\nInput limit inequalities (4.5) reduce to\na12 ) --~-434 (4.9)\nOutput limit inequalities (4.6) reduce to\na12 ~ ___a34 O.lO)",
+            "For a crank-rocker we satisfy (4.10)\nFor a rocker-crank we satisfy (4.9)\nFor a drag-link mechanism with change points at2 =a34.\n271\n(iv) Locking condith)ns\n(a) We note that mot ion is not possible if e = rr/2 since there can be no transmission.\n(b) Input locking can occur if $3 = 0 for\n0 3 = 0 or zr and from (4.3) the mechanism proport ions would need to satisfy\n$4.4. -- -4- [at22 - - (a23 --I- a3~ \" -- a4t)2] ~\" cosec\n(c) Locking of the output link can occur if\n$2 = 0 for 02 = 0 or n and from (4.4) the mechanism proport ions would need to satisfy.\nSt t -- _ [a34. 2 - ( a 4 t - a 2 3 Ta t2 )2 ]~cosec~\nThe R - C - R - C mechanism may be derived f rom the R - C - C - R mechanism by simply interchanging a revolute and cylindric pair. The classification is similar to the R - C - C - R mechanism.\nAcknowledgement--The authors wish to acknowledge the financial assistance of the Science Research Council (Grant B/SR/4354).\nR e f e r e n c e s [1] YANG A. T. and FREUDENS'rEtN F. Application of dual number quaternion algebra to the analysis of\nspatial mechanisms. J. AppL Mech, 31, No. 2; Trans. ASME (Series E) 86, (1964). [2] YANG A.T. Application of quaternion algebra and dual numbers to the analysis of spatial mechanisms.\nDoctoral Dissertation, Columbia University, New York (1963); University Microfilms, Ann Arbor, Michigan. [3] FREUDENS'I'Ett~ F. and SANOOR G. N. Kinematics of mechanisms. Mechanical Design and Systems Handbook. McGraw-Hill, New York (1964). [4] GILMAR'nN M. J. and DUFr\u00a2 J. A classification of four-link mechanisms. Departmental Research Report, No. ME-A68-2, Liverpool Regional College of Technology. [5] BENNE'r'r G. T. A new mechanism. Engineering 76 (1903). [6] HARmSBERGER L. and SONt A. H. A survey of three-dimensional mechanisms with one general con-\nstraint. ASME paper No. 66-Mech--44. [7] GtLMARTtN M. J. and DUFFV J. Limit positions of four-link spatial mechanisms--2. Mechanisms\nhaving revolute, cylindric and prismatic pairs. J. Mechanisms 4, 273-281 (this issue). [8] DXMENTBErtG F. M. and YOSLOVXCH I. V. A spatial four-link mechanism having two prismatic pairs.\nJ. Mechanisms 1, 291-300 (1966).\nAppendix ; Separation of the Locking Position Expressions for the R - C - C - C Mechanism into Dual Parts (i) Input locking position\nc\u00b0s 01[al 2 c\u00b0s~q 2 sin ~4t \"1- a41 sin=12 c0s=41] - S1 t sin=12 sin ~41 sin0t\n=(a23 -t-a3,t)sin(=23 +a3,~)--a12 sin~q2 cos=4t --a4t cos :q 2 sin=,, t (A.1)\nc\u00b0s 02[a12 c\u00b0s=t2 sin(~23 +=34) + (a23 ---a3,*)sin =12 cos(:t-, 3 ---+ ~3,,)]\n- $2 sin = t ~ sin(=23 + =34)sin 02 = a,t 1 sin ~,,t - at 2 sin at 2 c0s(=23 - =34)\n- (a23 _+a3,,)cos =12 sin(=23 +=3,*) (A.2)"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure2-1.png",
+        "caption": "Fig. 2. The PZL M28 Skytruck aircraft [9, 10]",
+        "texts": [
+            " D16AT-joint model is a smoothed model used in riveted joint analysis (point 4). D16AT-rivet model was used for the skin and the doubler elements in the rivet calculation. PA25-rivet model was used in the same analysis for rivet elements (point 5.1). D16AT properties were assumed according to USSR aerospace industrial branch standard OST 1 90070-92. PA25 properties were taken from [8]. Models used in riveting process simulation were described in point 5.2. The PZL M28 Skytruck (army version called Bryza, figure.2) is a two-engined, braced high-wing monoplane for passenger/cargo transportation purpose with metal structure. The maximum take-off weight is 7500 kg. The aircraft has short take-off and landing (STOL) capability and can operate from short, unprepared runways. It is manufactured by PZL Mielec, and it\u2019s used by e.g. Polish Army, Polish Border Guard as well as customers from USA, Venezuela, Columbia, Vietnam, Indonesia and Nepal. Jerzi Kaniowski et al. 944 It is a semi mono-cook metal structure. The airplane wing consists of a rectangle centre wing and two trapezium outer wing parts"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000875_jnm.673-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000875_jnm.673-Figure2-1.png",
+        "caption": "Figure 2. The two-port element implementing the coupling between the electric circuit and the equivalent magnetic circuit.",
+        "texts": [
+            " The topology of the magnetic equivalent circuit is directly derived from the magnetic structure and each winding is represented by an ideal voltage source of value Ni; where N is the number of turns and i is the winding current. Each flux tube is represented by an equivalent resistor whose resistance, possibly nonlinear, depends on the geometry and permeability of the magnetic material. A typical electrical equivalent of a magnetic structure is shown in Figure 1. The coupling between the electric circuit and the magnetic equivalent model is represented by a two-port element: its electrical port involves the v inductor branch voltage and the i current; the magnetic port involves F and F as shown in Figure 2. In particular we have F \u00bc Ni and v \u00bc NdF=dt: As an example by considering Figure 1 the independent voltage source connected to the A and B nodes is substituted by the two-port in Figure 2(b). Copyright # 2008 John Wiley & Sons, Ltd. Int. J. Numer. Model. 2008; 21:309\u2013334 DOI: 10.1002/jnm The simple model presented in the previous subsection can be enhanced by considering eddy currents. These contribute to F that therefore can no longer be assumed uniform along the cross section of each flux tube. Modeling of eddy currents requires a discretization of the core in N slices that suitably link flux [16, 17], inside which both magnetic and electric quantities are assumed to be uniform",
+            " The variation of Fk due to eddy currents in the kth flux tube can be related to the electromotive forces associated with the electric slices through the relation Rk 1ik 1 Rkik \u00bc dFk dt where Rk is the electrical resistance of the kth slice and ik is the eddy current. In this way, a set of N 1 equations is obtained; by solving it with respect to the currents, we have ik \u00bc 1 Rk Xk 1 l\u00bc1 dFl dt \u00f01\u00de By taking into account the circuit topology and flux linkage, a ladder network is obtained [16, 17]; the constitutive cell of the ladder network is shown in Figure 4. Note that the two controlled voltage sources model the relationship between the eddy current and the flux variation; it is similar to the two-port already shown in Figure 2. Furthermore, we Copyright # 2008 John Wiley & Sons, Ltd. Int. J. Numer. Model. 2008; 21:309\u2013334 DOI: 10.1002/jnm have ik \u00bc 1=Rk\u00f0d %Fk=dt\u00de; that is, the magneto-motive force across the controlled voltage source, labelled as A in Figure 4, is proportional to the derivative of its branch current %Fk: This means that it can be substituted by an inductor of 1=Rk value. However, we prefer the representation shown in Figure 4 since there is more evidence that eddy currents introduce power losses through the Rk resistors"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000762_tsmcb.2007.909943-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000762_tsmcb.2007.909943-Figure1-1.png",
+        "caption": "Fig. 1. Object-handling task.",
+        "texts": [
+            " Generally, by selecting proper system parameters, the spheres can be picked up by the probe due to the attractive force between them [1]. On the other hand, the job of releasing the spheres needs totally different techniques. Various placing methods have been introduced in [1], [13], [14], and [31]. For instance, in [31], electrostatic interaction is utilized. In this paper, we will concentrate our work on the pickup task of microspheres. For the purpose of designing a controller for the objecthandling task, we shall restrict ourselves with the intricacies of the physics during the pickup process, as shown in Fig. 1. The adhesion forces are dominant in the system. They are considered to play an important role in the manipulation process. These are given by the following: \u2022 van der Waals forces; \u2022 surface tension (or capillary); \u2022 electrostatic (or Coulomb) forces. A. Van der Waals Force The van der Waals force acts between atoms resulting from the interaction between electrons in the outermost bands rotating around the nucleus of the atoms. An overview of it is given in [19]. Van der Waals forces are present in every environmental condition"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001257_j.euromechsol.2008.11.006-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001257_j.euromechsol.2008.11.006-Figure2-1.png",
+        "caption": "Fig. 2. Differential element of the tube.",
+        "texts": [
+            " The membrane is filled with two fluids, \u03a91 and \u03a92 and the interface of two internal fluids is \u0393 . This system is also surrounded by two fluids, \u03a93 and \u03a94. The membrane profile can be defined by the parametric equations X(S) and Y (S) where S is the arc length along \u2202\u03a9 measured from some arbitrary reference point. The governing differential equations of a pliable tube in terms of the applied internal and external pressures can be obtained by considering the free body diagram of a differential element from S to S + dS as depicted in Fig. 2. The geometrical relationships are\u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 dX dS = cos\u03c8, dY dS = sin \u03c8, S \u2208 \u2202\u03a92, (1) At the general position S , the tension in the tube is N and the tube element makes an angle \u03c8 with the horizontal axis. At the section S + dS the tension is N + dN and the angle is \u03c8 + d\u03c8 . The equilibrium equations in horizontal and vertical directions are (N + dN) cos(\u03c8 + d\u03c8) \u2212 N cos\u03c8 + P dS sin ( \u03c8 + d\u03c8 2 ) = 0, (2a) (N + dN) sin(\u03c8 + d\u03c8) \u2212 N sin\u03c8 \u2212 P dS cos ( \u03c8 + d\u03c8 2 ) \u2212 \u03bbg dS = 0, (2b) Expanding and simplifying the equations give dN dS \u2212 \u03bbg sin\u03c8 = 0, (3a) N d\u03c8 dS \u2212 P \u2212 \u03bbg cos\u03c8 = 0, (3b) where P is transmural pressure (Paidoussis, 2004) which is the difference between internal and external pressures P int \u2212 Pout and \u03bb is the mass per unit length of the tube"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003668_tcst.2011.2167974-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003668_tcst.2011.2167974-Figure1-1.png",
+        "caption": "Fig. 1. Model coordinates , and controls , , . Descent angle and vertical velocity shown here negative during descent.",
+        "texts": [
+            " The proposed control scheme is presented in Section IV and its formal properties are analyzed. Section VI provides simulation results while Section VII concludes our work. Since this work is aimed specifically at future ATM, the model of each agent is chosen to closely capture the motion of civilian fixed-wing aircraft. Thus, each agent is considered as a unicycle on the horizontal plane, that can also use its vertical linear velocity to adjust its altitude (1) where is the projection of the agent\u2019s position on the horizontal plane, see Fig. 1. Moreover, , where is the heading angle, i.e., the angle between the agent\u2019s longitudinal axis and the global axis, and is the altitude. The control vector comprises the horizontal, vertical, and angular velocities , , and , respectively. We use the climb or descent angle defined by the resultant velocity vector and the horizontal plane, , with representing climbing. Compared to a unicycle-like model with the linear speed and the three angular velocities as inputs (like the one used in [4]), model (1) decouples horizontal and vertical manoeuvring, allowing independent regulation of the vertical velocity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure7-1.png",
+        "caption": "Fig. 7. The 3D-holospectrum of the original unbalance response.",
+        "texts": [
+            " Since the major axis always has the maximum precession value, the reduction of its magnitude surely indicates the improvement of the rotor balance. Fig. 6 shows a sketch map of a 300 MW turbo generator set. The whole machine set consists of one high pressure cylinder (HP), one intermediate pressure cylinder (IP), two low pressure cylinders (LP1 and LP2)and one generator (GEN). All these machines are connected by rigid couplings. The first startup after an overhaul showed that the vibrations detected at bearing section No. 1 and No. 3 exceeded the alarm level. The largest peak-to-peak value was even over 200 lm. Fig. 7 gives the 3D-holospectrum of the original response. Since the rest part of the machine set ran very well and the trial responses were also insignificant, only the vibrations of the first four sections were considered here. The vibration signals were carefully analyzed and mass unbalance was confirmed as the main source of vibration. Since the balancing history of this rotor was not available, its correction was actually a trial and error procedure. Totally three balancing planes were used, which were set on the bearing flange close to section No"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000239_jctb.5020230709-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000239_jctb.5020230709-Figure1-1.png",
+        "caption": "Figure 1 . Encapsulated spherical enzyme particle.",
+        "texts": [
+            " Specifically, the transient reactor behaviour of encapsulated CSTR and packed bed tubular reactor systems was simulated in this study. A combined rate control model of reaction and membrane diffusion was employed to express the kinetics of the substrate consumption catalysed by the encapsulated enzyme particle. Effects of various parameters such as the Peclet number, Michaelis constant, catalytic reaction constant, and membrane diffusional resistance on the dynamic substrate concentration changes at various positions of the reactor systems were examined. The diagram of an encapsulated spherical enzyme particle is shown in Figure 1. Assumptions made in deriving the model equations are: (3) The membrane with which the enzyme is occluded is inert. (b) No hindrance effect on substrate transport through the pore space of the membrane exists. (c) No conformational change takes place in the enzyme's active site during encapsulation. (d) All rate processes occur under an isothermal condition. (e) The kinetic rate constant and the Michaelis constant of the encapsulated enzyme are identical with those of the free native enzyme. Encapsulated enzyme reactor systems 533 (f) The rate processes are under steady state conditions at any moment"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001907_1.3516354-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001907_1.3516354-Figure1-1.png",
+        "caption": "FIGURE 1. Schematic diagram of a 3 DOF spatial manipulator",
+        "texts": [
+            " Tian et al [4] introduced obstacles in the manipulator\u2019s workspace and calculated the optimum trajectory that describes minimum torque as well as avoids collision with the obstacle. This work aims at an energy optimal trajectory planning of 3 degrees of freedom (DOF) spatial manipulator in the presence of an obstacle in its workspace. The optimization required here consists of two objectives; viz. energy minimization of the actuators of the manipulator and collision avoidance with an obstacle inside the manipulator\u2019s reachable workspace. Fig.1 shows the schematic diagram of 3 DOF spatial robotic manipulator considered in the present work. Dynamics equations are necessary to formulate the forces/torques required to produce the desired motion of the manipulator. The link coordinate frame is attached as per Denavit-Hartenberg conventions [5] to the distal joint of each link to describe the spatial displacement between neighboring link coordinate frames to obtain the kinematic information. For deriving dynamics equations and finding the torque variation over the movement cycle, Lagrange-Euler formulation has been used"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.5-1.png",
+        "caption": "Fig. 3.5. Anthropomorphic arm at an elbow singularity",
+        "texts": [
+            " Arm singularities are characteristic of a specific manipulator structure; to illustrate their determination, consider the anthropomorphic arm (Fig. 2.23), whose Jacobian for the linear velocity part is given by (3.38). Its determinant is det(JP ) = \u2212a2a3s3(a2c2 + a3c23). Like in the case of the planar arm of Example 3.2, the determinant does not depend on the first joint variable. For a2, a3 = 0, the determinant vanishes if s3 = 0 and/or (a2c2 + a3c23) = 0. The first situation occurs whenever \u03d13 = 0 \u03d13 = \u03c0 meaning that the elbow is outstretched (Fig. 3.5) or retracted, and is termed elbow singularity . Notice that this type of singularity is conceptually equivalent to the singularity found for the two-link planar arm. By recalling the direct kinematics equation in (2.66), it can be observed that the second situation occurs when the wrist point lies on axis z0 (Fig. 3.6); it is thus characterized by px = py = 0 and is termed shoulder singularity . Notice that the whole axis z0 describes a continuum of singular configurations; a rotation of \u03d11 does not cause any translation of the wrist position (the first column of JP is always null at a shoulder singularity), and then the inverse kinematics equation admits infinite solutions; moreover, motions starting from the singular configuration that take the wrist along the z1 direction are not allowed (see point b) above)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure7-1.png",
+        "caption": "Fig. 7 Load distributions on teeth: (a) standard tooth, (b) modified tooth",
+        "texts": [],
+        "surrounding_texts": [
+            "3.1 Force distribution on a standard tooth profile In order for gears to be continuously in mesh, a pair of meshing gear teeth should not break their contact before another pair of teeth behind them starts to mesh; see Fig. 4. That is, two gear teeth pairs should be in contact at the same instant, which helps with dividing the total tooth load into two parts; see Fig. 5. Therefore, load distribution along the meshing line is not constant, but changes step by step, and force distribution along the profile depends on contact ratio. As the contact ratio approaches 1, the ratio of the length of a single meshing area to that of a double meshing area increases, as illustrated in Fig. 5(b). 3.2 Procedure of width modification With this type of modification, tooth width, b, has been widened in order to maintain a constant ratio of gear load to gear width (F/b), which tends to vary continuously along the profile of a standard gear. In Fig. 6, a normal and a modified gear model are shown along with their respective load distributions in Figs. 7 and 8. Face width, b, is not standardized, but generally, 9 m < b < 14 m. The wider the face width, the more difficult it is to manufacture and mount the gears so that contact is uniform across the full face width [20]. However, where the volume is limited (for instance, in gearboxes), then recommended values lie in the range 6 m < b < 12 m. If the given values are considered, then width limits in gear width modification will remain within acceptable standards. 3.3 Manufacturing methods Nowadays, the needs of high speed and high load carrying capacities of gears have significantly increased. On the other hand, in the manufacturing of gears, quick production and low expenses have become the main criteria. The emergence of new heat-resistant and high strength plastic materials has initiated improvements in gear manufacturing methods. This includes production of gears with the die method. In this method, molten or powdered material is poured into a specially designed mould having the shape of the final required gear. Examples of die forming processes are injection moulding, the powdered metal process, and forging [19]. With the above mentioned manufacturing methods, it is possible to manufacture gears whose teeth widths have been modified. In cutting methods, a gear is machined on a turning machine, as seen in Fig. 9, in a way resembling the load distribution on the gear surface, and then gear teeth are cut with gear teeth cutting methods. In the injection method, the gear is obtained with the help of dies, as explained before. 3.4 Hertzian stress along contact path Gear teeth are subjected to Hertz contact stresses and various surface damages. Overloading is one of the basic factors that cause abrasion, and scoring. Generally, a good correlation has been observed between spur gear surface fatigue failure and the computed elastic surface stress (Hertz stress) [20]. As shown in Fig. 10, the radius of the pinion at the initial contact point A is \u03c11 and that of the gear is \u03c12. For external spur gears, the Hertzian contact stress at point A is \u03c3H = \u221a \u221a \u221a \u221a \u221a \u221a F ( 1 \u03c11 + 1 \u03c12 ) \u03c0b ( 1\u2212\u03bd2 1 E1 + 1\u2212\u03bd2 2 E2 ) (1) where F is normal force, b is the gear tooth face width, E is Young\u2019s modulus, and \u03bd is the Poisson ratio. One of the fundamental relationships is evident from this equation. Because of the increased contact area with load, stress increases only as the square root of load F (or square root of load per mm of face width, F/b). As seen in Fig. 8, although the F/b ratio in a standard gear has changed along the meshing line; that of a modified gear has remained almost linear along the line and the value is halved. As a result, as this value drops to almost a half, the Hertz stresses will also drop in proportion to the square root of the F/b ratio. Moreover, with the linear distribution in modified gears, especially along single and double meshing zones, the instantaneous changes in the F/b ratio can be eliminated from the damaging effects. Based on the fact that the value of loading in single meshing zone is twice that at the double mesh, the twofold increase in gear tooth width, b, in the single meshing zone will cause a corresponding decrease in contact pressure at the rate of around 40 to 50% as calculated from (1). That is, the contact pressure values between points DB will fall by 40 to 50%. Also, the shaded area in the contact pressure diagram (Fig. 11) will disappear. 3.5 Temperature distribution along the contact path Metallic gears generally operate in oily media, whereas plastic gears run dry. In this respect, metallic gears have the advantage of dissipating heat out of the gear teeth. Plastic gears have no such advantage, and due to their lower thermal conductivities, the gears are subject to damages as a result of tooth temperatures. A large portion of the heat produced in gear teeth is due to friction and the transmitted load per unit tooth width is directly proportional to the friction-based heat produced. This shows that this heat has direct effects on load sharing. Flash temperatures of gears can be calculated with the following expression: [21]. Tf = 1.11\u03bcF \u221a W1 \u2212 \u221a W2 b \u221a 2k\u03c1ca (2) Here, \u03bc represents the friction coefficient, F the normal load, W1 and W2 indicate surface velocities, b is the gear tooth width, k represents the thermal conductivity, \u03c1 stands for density, c is specific heat capacity, and a is the semi-width of contact. The semi-width of contact is calculated from the following expression a = \u221a 4FR b\u03c0E (3) Here, E is Young\u2019s Modulus and R is the equivalent radius of curvature of a meshing point which is calculated from the following expression. 1 R = 1 \u03c11 + 1 \u03c12 (4) where, \u03c11 and \u03c12 are radii of contact point curvatures in the pinion and gear. Figure 12 shows sliding speed and load distribution in standard gears and theoretical flash temperature distributions. In (2), apart from gear material and tooth width, if all other geometrical properties are kept constant and the tooth width in the single meshing zone is doubled, then a decrease of 40 to 50% in flash temperatures will be observed in this area. That is, flash temperature values between points DC and CB will fall by 40 to 50%. Also, the shaded area in the flash temperature diagram (Fig. 12) will disappear. Dissipation of heat from gears, apart from conditions like surrounding medium temperature and thermal conductivity, is also closely dependent on gear tooth surface area. In addition, with this modification, the amount of heat dissipated will increase and the modification will help to create equilibrium between heat generation and heat dissipation. In measurements made on gears given in Fig. 6, an increase of 20% in the total surface areas of modified gears was found. Therefore, as seen in Fig. 13, this area increase has positive effects in dissipating heat away from the gear tooth."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000293_tmag.2006.871428-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000293_tmag.2006.871428-Figure7-1.png",
+        "caption": "Fig. 7. Detail of (a) mesh 1 and (b) mesh 2. Slipping interface for LM and Maxwell stress tensor integration surface S.",
+        "texts": [
+            " If the interpolation order is augmented to two (using hierarchic interpolation on and hierarchic complete elements on moving band), the results presented in Fig. 5 are obtained. In this case, against expectations, we obtain a worst result with more oscillations than with first-order interpolation. Because the torque calculation in thin air gaps may surprise, the switched reluctance motor [4] shown in Fig. 6, which has an air gap of 0.25 mm, is also simulated. For this machine, two discretizations (Fig. 7) are used: mesh 1 (with 2486 elements, 1605 nodes, 120 edges on and four layers of elements in the air gap) and mesh 2 (with 4861 elements, 2730 nodes, 240 edges on , and 2 layers of elements in the air gap). The air-gap thickness g is 0.25 mm. The interface is placed in the middle of the air gap. The MB is placed in the layer next to the rotor side of . The torque is always calculated in the layer of quadrilateral elements placed next to the stator side of . The Maxwell stress tensor contour passes by the midpoints of the quadrilateral element edges (Fig. 7). The rotor displacement step is 0.15 for all results. To overpass one edge on 10 (for mesh 1) and 5 (for mesh 2), rotor displacement steps are necessary. If we compare conforming and nonconforming formulations for mesh 1 with first-order interpolation, Fig. 8 shows that the LM gives better results than the MB. However, Fig. 9 shows strong oscillations for the torque calculation with second-order interpolation, with either LM or MB, as observed for the first example. The results presented in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000534_j.jsg.2008.01.010-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000534_j.jsg.2008.01.010-Figure5-1.png",
+        "caption": "Fig. 5. (a) A Class 3 fold. (b) Class 1B fold obtained by interchanging the outer and inner arcs of the Class 3 fold shown in (a). (c, d) Class 3C and Class 3A folds, respectively. The isogon rosette and the characteristic curve are shown below each fold. The characteristic curve is a rectangular hyperbola, circle, and hyperbolas in (a), (b), (c) and (d), respectively. A1 and A2 are asymptotes of hyperbola, and a and b are lengths of the transverse and the conjugate axes of hyperbola. The axial ratio of the hyperbola is given by the aspect ratio of the rectangles. Ellipses inside the rectangles are strain ellipses. Rs, Two-dimensional strain ratio.",
+        "texts": [
+            " The estimation of flattening strain for Class 3 folds is based on the assumption that the buckling of a multilayer sequence, consisting of an incompetent layer sandwiched between two competent layers, produces a Class 3B fold in the incompetent layer. The existing definition of a Class 3B fold (Zagor cev, 1993) requires its association with a Class 1B fold, such that the layer thicknesses along the axial trace of the Class 1B and the Class 3B folds are same. Based on this definition, we give a simple geometrical criterion for classification of Class 3 folds. Consider a Class 3 fold with h1 and h2 as the hinge points located on the inner and the outer arcs, respectively (Fig. 5a). While keeping the outer arc fixed, move the inner arc upwards along the axial trace until the hinge point h1 shifts to the point h3, such that h1h2 \u00bc h2h3 (Fig. 5b). This interchange of the outer and inner arcs transforms the Class 3 fold into a Class 1 fold (Fig. 5b). If the geometries of the transformed folds are Class 1A, 1B or 1C this implies that the original Class 3 fold is a Class 3A, 3B or 3C fold, respectively. The superimposition of a homogeneous flattening strain on a Class 3B fold at an angle q equal to 0 or 90 modifies its geometry into a Class 3C or 3A fold, respectively (Fig. 5c,d). For a Class 3B fold, the curve through the end points of the isogon rosette is characteristically a rectangular hyperbola of unit axial ratio, y2/a2 x2/a2 \u00bc 1, such that the centre of hyperbola coincides with the common point of intersection of the isogons, with its transverse axis parallel to the fold axial trace (Fig. 5a). As the flattening modifies the Class 3B fold into a Class 3A or Class 3C fold, the rectangular hyperbola transforms into a hyperbola, y2/a2 x2/b2 \u00bc 1 (a s b), which has two non-orthogonal asymptotes (Fig. 5c,d). The axial ratio of the non-rectangular hyperbola obtained from the isogon rosette in the Class 3A or Class 3C folds directly gives the strain suffered by the original Class 3B fold during the process of flattening. The maximum principal strain parallels the transverse axis of the hyperbola for Class 3A folds, whereas it parallels the conjugate axis of the hyperbola for Class 3C folds (Fig. 5c,d). In all obliquely flattened Class 3B folds, the major axis of the strain ellipse is inclined with respect to the transverse axis of the hyperbola, i. e., d the \u2018\u2018isogon rosette\u2019\u2019 method (Table 1). Types 1 to 3 are defined in the text. Rs, strain ratio; q, angle between the maximum principal strain and fold axial trace; RMS, root mean square error in the best-fit ellipse obtained by algebraic fitting. Source for folds: 1 and 2, Hudleston (1973b); 3e6, Dietrich (1969); 7e 26, unpublished; 27, Ding and James (1985); 28e33, unpublished"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002323_978-3-642-25486-4_25-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002323_978-3-642-25486-4_25-Figure7-1.png",
+        "caption": "Fig. 7. Elasticity contribu",
+        "texts": [],
+        "surrounding_texts": [
+            "1. Merlet, J.-P.: Parallel Robo 2. Stewart, D.: A Platform Engineers Proceedings, Vo 3. Clavel, R.: Device for the No. 4976582 (December 1 4. Clavel, R.: Delta, a fast ro Symposium on Industrial R 5. Tsai, L.-W.: Robot Analy Wiley & Sons (1999) 978-0 6. Nefzi, M.: Kinematics an Vector-based kinematic cal 7. Gosselin, C.: Stiffness Ma and Automation 6(3), 377\u2013 8. Timoshenko, S.: Strength & Distributors (2004) 978- Stiffness Analysis of Clavel\u2019s DELTA Robot to maximum deflection for reference load in X- and Y-directi tors to maximum deflection for reference load in Z-direction ts. Springer, Heidelberg (2006), 978-1402041327 with Six Degrees of Freedom, UK Institution of Mechan l. 180(15), Pt. 1 (1965) Movement and Positioning of an Element in Space, US Pa 1, 1990) bot with parallel geometry. In: Proceedings of 18th Internatio obots, pp. 91\u2013100 (1988) sis: The Mechanics of Serial and Parallel Manipulators. J 471325932 d Dynamics of Robots, Exercise RWTH Aachen, Germa culations for parallel robots, not published (2010) pping for Parallel Manipulators. IEEETransactions on Robo 382 (1990) of Materials: Elementary Theory and Problems. CBS Publish 8123910307 249 on ical tent nal ohn ny: tics ers"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003956_ejc.18.411-421-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003956_ejc.18.411-421-Figure2-1.png",
+        "caption": "Fig. 2. Schematic picture of the transmission system.",
+        "texts": [
+            ") The remaining alternative is that the normalized prediction error tends to zero while the prediction error itself tends to infinity. It cannot be ruled out that the parameter \u03b8 then could become unbounded or approach a limit cycle as time tends to infinity. Supported by the analysis results from the most recent sections, the remainder exemplifies how to apply the design procedure from Section 2 through two design problems inspired by [8] and [2]. A schematic picture of a flexible transmission system is shown in Fig. 2. It consists of three horizontal pulleys connected by elastic belts. The first pulley is controlled by an electric motor with local feedback, using the reference for the first pulley as its input. The goal is to control the position of the third pulley, where a PC is used to control the system. The plant for three modes of operation is described in discrete time as No load: P(q\u22121) = 0.28q\u22122 (1 + 0.18q\u22121 + 0.93q\u22122) \u00d7 (q\u22121 + 1.8q\u22122) (1 \u2212 1.6q\u22121 + 0.95q\u22122) (8a) Half load: P(q\u22121) = 0.1q\u22122 (1 \u2212 0.17q\u22121 + 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure3-1.png",
+        "caption": "Fig. 3. Substructure A+B: constrained BIW for torsional stiffness",
+        "texts": [
+            " (1) Equation (1) indicates how the system matrix of the coupled system HC can be calculated from the system matrices of the components (HA and HB) The indices are related to input, output and coupling points as indicated in Fig. 1. The FBS technique can be used to convert the free-free system to a constrained system: \u2022 subsystem A: free-free FRF's of BIW as shown in Fig. 2 \u2022 subsystem B: ground To represent the static test bench condition for torsional stiffness, the rear domes will be grounded (Fig. 3). To represent the static test bench condition for bending stiffness, rear and front domes will be grounded (Fig. 4). Forced Response The bending stiffness of a vehicle body is measured by clamping the body at the four domes and applying a load at the 4 seat bolting positions (F1 to F4) as is represented in Fig. 5. (2) After calculating the coupled system matrix Hc from equation (1) the displacement in the output points d(\u03c9) can be easily calculated using a forced response described in equation (2)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003542_j.elstat.2012.04.003-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003542_j.elstat.2012.04.003-Figure10-1.png",
+        "caption": "Fig. 10. The dimensions of the Bell 206 helicopter.",
+        "texts": [
+            " The greater discrepancies near the sphere where the field increases rapidly are perhaps because any small positioning errors of the meter up or down, or sideways, or away from the spherewill cause a decrease in the measured field. Possibly the effect of the finite size of the metal parts of the electric field meter itself may lower the reading in a non-linear and rapidly-changing field. 4.3. Electric field in the presence of a floating-potential helicopter model Since the wing, skid, tail and other parts with small radii of curvature will have a high field close to them; these parts were accurately modeled in the finite-element calculations, in order to obtain reliable results, using the dimensions and layout shown in Fig. 10. The fuselage, and other parts which affect the electric field less, were simplified in the model used for the computation, as shown in Fig. 11a; and similarly in the model built for the experiments (and shown in Fig. 11b). The diameter of the cylindrical cabin was 0.19 m. The measurement of the field in the presence of the helicopter was carried out in the same manner as with the sphere. Again the voltage applied to the line was 200 kV.The measurement line extended from the transmission line in the normal direction and horizontally towards the rotor blades"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001171_demped.2007.4393076-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001171_demped.2007.4393076-Figure1-1.png",
+        "caption": "Fig. 1. Traces of current change phasor resulting from different static air gap asymmetries combinations of voltage pulses in the complex plane (upper);",
+        "texts": [
+            " Thus seven different voltages (six active, On a lI and the two zero voltage phasors) can be applied, value of caused by the inherent asymmetries is only a few * The voltage drop of the stator resistance. percent of that of Imean leading to a direction of the current * The influence of the back emf caused by the time slope in the complex plane that is dominated by the spatial derivative of the rotor flux. direction of the voltage pulses applied (vsi - ySN,). This is * And finally the value of the transient leakage inductance 1 shown in Fig. 1 for the different combinations of inverter itself. switching states. As the pulses are applied only for a very short duration The upper diagram in Fig. 1 shows the traces of the current (some 1 Otis) the dc link voltage as well as the time derivative change in the stator fixed reference frame resulting from the of the rotor flux can be considered constant during the short different combinations of two non-parallel active voltage time periods necessary for the detection. Their influence can pulses. The machine was operated without fundamental wave thus be eliminated by taking samples of the current slope Of current. Only the pulse sequences were injected"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002285_660432-Figure15-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002285_660432-Figure15-1.png",
+        "caption": "Fig. 15 - Standard cap screw",
+        "texts": [
+            " 10 It is believed that continued use of the oscillating torque type of testing machine will be valuable in defining the chat, acteristics of self-locking bolt configurations. However, obtaining a complete spectrum or map for a particular bolt is a lengthy process. It may be much better to use the stress level at which a bolt is to operate and to conduct one or a series of tests at this stress level in order to determine the comparative holding power of a number of designs. As an illustration of the data that can be obtained in this way, three different designs of 1/2-13-1 machine bolt were selected for testing. These are shown in detail in Fig. 15 (a standard cap screw), Fig. 16 (a self-locking bolt with a flexing head), and Fig. 17 (a new asymmetric bolt design using controlled storage of bending energy to obtain self- locking characteristics). All tests were run using '15,000 psi bolt shank stress as a starting point. The mean number of cycles to loosen is given in Table 1 along with the standard deviation of each mean, as obtained from a large number of tests on each design. This is seen to be a convenient method of rating bolts, since the data used in making Table 1 can be obtained rather quickly"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000647_tmag.2007.893803-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000647_tmag.2007.893803-Figure1-1.png",
+        "caption": "Fig. 1. Two-dimensional model in polar coordinates system.",
+        "texts": [
+            " In particular, analytical technique using the electromagnetic transfer relations theorem (TRT) that proposed by Melcher is more useful than the existing space harmonic method because it reduces analytical burden such as derivation of the governing equation [1], [2]. This paper presents an analytical solution to predict magnetic field distribution and to calculate parameters of permanent magnet (PM) synchronous motor equipped with surface-mounted magnets by using TRT in terms of 2-D model in polar coordinates system. The analytical results are validated by comparison with finite element analyses and experimental results. Fig. 1 shows the 2-D model of a typical 4-pole, 3-phase PM synchronous motor equipped with a parallel magnetized PM rotor and a slotted stator core. The main design specifications of PM synchronous motor are given in Table I. In order to establish analytical solutions for the magnetic field distribution, this paper assumes that the relative permeability of the stator core and rotor shaft is infinite and the current is distributed in an infinitesimal thin sheet at . Fig. 2 shows the simplified analytical model with only PM field for deriving transfer relations at the magnet surface [(d) and (e) in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000978_20090712-4-tr-2008.00015-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000978_20090712-4-tr-2008.00015-Figure8-1.png",
+        "caption": "Fig. 8. Level curves of Pd (x) (global convergence)",
+        "texts": [],
+        "surrounding_texts": [
+            "The control law given in (14) requires the complete state feedback. Moreover it also requires the knowledge of the temperature dependence of the kinetic coefficient k (T ) that appears in the expression of WT\u2217 (T ). In most practical applications, the kinetic coefficient is determined experimentally and the on-line measurement of the concentration is not always achievable. In this section we shall analyze the effect on the closed-loop convergence and stability of modeling errors in the kinetic coefficient k (T ). Let us assume that there is a modeling error on the kinetic coefficient. The aim is to stabilize the state (n\u2217 A, T \u2217). The controller is designed using the function k\u0302 (T ) instead of the real kinetic coefficient k (T ): k\u0302 (T ) = (1 + \u03be (T )) k (T ) (17) with \u03be (T ) > \u22121. Assumption 1. Despite the error on k (T ), the equilibrium value of nA is known: n\u2217 A = neq A (T \u2217) = \u03b4 Cin A V \u03b4 + k (T ) = \u03b4 Cin A V \u03b4 + k\u0302 (T ) Assumption 2. Despite the error on k (T ), the equilibrium value of the control input u\u2217 is known: u\u2217 c =\u2212\u03b4 (Tin \u2212 T \u2217) \u2212 \u03b3k (T \u2217)n\u2217 A =\u2212\u03b4 (Tin \u2212 T \u2217) \u2212 \u03b3k\u0302 (T \u2217)n\u2217 A Assumption 3. The control input has been designed such that, based on the estimated value of the kinetic coefficient, the desired equilibrium is asymptotically stable. From (14), the control input applied to the system is given by the following expression: uc (Tw \u2212 T ) = \u03bc ( n\u2217 A \u2212 nA \u2212 \u222b T T\u2217 w\u0302 (\u03c4) d\u03c4 ) + u\u2217 c (Tw \u2212 T \u2217) where we have introduced the following notation, for the sake of clarity: w\u0302 (T ) = w ( k\u0302 (T ) ) . Assumptions 1 and 2 imply that (n\u2217 A, T \u2217) is still an equilibrium of the closedloop system. Assumption 3 implies that the function w (y) and the parameter \u03bc have been chosen such that the following matrix is negative definite: \u039b\u0302 = \u239b \u239c\u239c\u239c\u239c\u239d \u2212 ( \u03b4 + k\u0302\u2217 ) \u2212n\u2217 A dk\u0302 dT \u2223\u2223\u2223\u2223\u2223 T\u2217 \u03b3k\u0302\u2217 \u2212 \u03bc \u2212\u03b4 + \u03b3n\u2217 A dk\u0302 dT \u2223\u2223\u2223\u2223\u2223 T\u2217 \u2212 \u03bcw\u0302 (T ) \u239e \u239f\u239f\u239f\u239f\u23a0 \u039b\u0302 is the matrix of the linearized system around the desired equilibrium state if the kinetics was indeed equal to k\u0302 (T ). As a consequence we have: tr \u039b\u0302 < 0 and det \u039b\u0302 > 0 The actual matrix of the linearized system around (n\u2217 A, T \u2217) is written as follows: \u039b = \u239b \u239c\u239c\u239d \u2212 (\u03b4 + k\u2217) \u2212n\u2217 A dk dT \u2223\u2223\u2223\u2223 T\u2217 \u03b3k\u2217 \u2212 \u03bc \u2212\u03b4 + \u03b3n\u2217 A dk dT \u2223\u2223\u2223\u2223 T\u2217 \u2212 \u03bcw\u0302 (T ) \u239e \u239f\u239f\u23a0 The trace and the determinant of \u039b are given by the following relations: tr \u039b = \u03a8\u2217 \u2212 \u03b4 \u2212 \u03bcw\u0302 (T \u2217) det\u039b =\u2212\u03b4\u03a8\u2217 + \u03bc [ (\u03b4 + k\u2217) w\u0302 (T \u2217) \u2212 nA dk dT \u2223\u2223\u2223\u2223 T\u2217 ] with \u03a8\u2217 = \u2212 (k\u2217 + \u03b4) + dk dT \u2223\u2223\u2223\u2223 T\u2217 \u03b3n\u2217 A (18) Using (17), we can define the following quantity \u03a8\u0302\u2217 by analogy with (18): \u03a8\u0302\u2217 = \u03b3n\u2217 A dk\u0302 dT \u2223\u2223\u2223\u2223\u2223 T\u2217 \u2212 ( k\u0302\u2217 + \u03b4 ) = \u03a8\u2217 (1 + \u03be\u2217) + \u03b3n\u2217 Ak \u2217 d\u03be dT \u2223\u2223\u2223\u2223 T\u2217 + \u03b4\u03be\u2217 where \u03be\u2217 = \u03be (T \u2217). As a consequence the trace and the determinant of \u039b(cl) can be rewritten as follows: (1 + \u03be\u2217) tr \u039b = \u03a8\u0302\u2217 \u2212 \u03b4 \u2212 \u03bcw\u0302 (T \u2217)\ufe38 \ufe37\ufe37 \ufe38 tr \u039b\u0302(cl)<0 \u2212 [ d\u03be dT \u2223\u2223\u2223\u2223 T\u2217 \u03b3k\u2217n\u2217 A + 2\u03be\u2217\u03b4 + \u03bc\u03be\u2217w\u0302 (T \u2217) ] By applying the same development on the determinant, the following expression is obtained: (1 + \u03be\u2217) det \u039b = det \u039b\u0302 + \u03be\u2217\u03b4 (\u03b4 + \u03bcw\u0302 (T \u2217)) + n\u2217 Ak \u2217 d\u03be dT \u2223\u2223\u2223\u2223 T\u2217 (\u03b3\u03b4 + \u03bc) The closed-loop equilibrium is asymptotically stable if and only if the trace is strictly negative and the determinant is strictly positive. Using the inequalities of Assumption 3, this means that if \u03be\u2217 > 0 and d\u03be dT \u2223\u2223\u2223 T\u2217 \u2265 0, the closed-loop equilibrium is asymptotically stable. But if these two conditions are not fulfilled, then the closedloop equilibrium can become unstable. Nevertheless, if the function w (y) has been taken sufficiently large such that: (\u03b4 + k\u2217) w\u0302 (T \u2217) \u2212 nA dk dT \u2223\u2223\u2223\u2223 T\u2217 > 0 then det \u039b(cl) is increasing with \u03bc such that there is a lower bound on \u03bc that ensures the local asymptotic convergence (see Figure 9)."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002017_10426914.2010.496126-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002017_10426914.2010.496126-Figure11-1.png",
+        "caption": "Figure 11.\u2014Relative density contour of the cylindrical specimen with a bag after cold isostatic pressing (the bag is not displayed).",
+        "texts": [
+            "\u2014von Mises stress contour of the cylindrical specimen with a bag after cold isostatic pressing (the bag is not displayed). Figure 10.\u2014Relative density contour of the cylindrical specimen without a bag after cold isostatic pressing. Figure 12.\u2014Comparison of experimental results and simulation results of relative density under different pressures. right corner. The contours of relative density with a bag and without a bag are shown in Figs. 10 and 11. The difference of relative density of the whole model in Fig. 10 is small; that is, 0 7672 \u2212 0 7677 = \u22120 0005. The contour of the relative density in Fig. 11 is similar to that of von Mises stress. However, because of the corner effects the relative density of the upper right corner is a little smaller than that of the main area; that is, from 0.76 to 0.77. From Fig. 11 only a small part of the upper right corner has the lower density region. The density of most regions is uniform. Simulation results with and without a bag all are close to the experimental results of 0.74. A comparison D ow nl oa de d by [ U ni ve rs ity o f T as m an ia ] at 1 7: 01 0 1 Se pt em be r 20 14 Figure 14.\u2014Dimensions of the cross section of the turbine component (unit:mm). Figure 15.\u2014Turbine component after selective laser sintering. of experimental results and simulation results of relative density under different pressures shows a similar tendency, which is illustrated in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002937_j.jlumin.2011.03.038-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002937_j.jlumin.2011.03.038-Figure2-1.png",
+        "caption": "Fig. 2. Schematic diagram of the FI cell.",
+        "texts": [
+            " They had the following drawbacks: (1) large dead volume; especially when RE was placed near the inlet, it would reduce both the detection sensitivity and separation efficiency of sample; (2) high IR drop; especially when RE was located at the downstream, if caused high overpotential and thus, decreased the sensitivity of ECL detection and (3) high flow resistance made it difficult to remove away possible gas bubbles, which would cause significant noise [20]. In order to overcome these shortcomings, in this paper, a FI cell (Fig. 2) was designed, which consists of an inlet, an outlet, an Ag/AgCl RE, a Pt CE and a glassy carbon WE, forming a sample champer of 40 mL. The inlet was near the three electrodes, which were located near each other (the distance between the three electrodes was 2 mm), resulting in a smaller dead volume and a lower IR drop in comparison with the conventional ones, which improve the detection sensitivity. Meanwhile, the positions of inlet, outlet and electrodes formed a lower flow resistance, which made it difficult for the electrochemically generated gas to accumulate in the thin layer"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure10-1.png",
+        "caption": "FIG. 10. Bearing bushes for profile grinder spindle.",
+        "texts": [
+            " Grinding Machine Spindles 463 To ensure long life the bushes in which the hardened steel balls reciprocate were made from accurately lapped tungsten carbide. The clearance being the minimum required to prevent excessive leakage while maintaining adequate lubrication. Conical Hydrostatic Bearings The hydraulic power source used to drive the motor can also be employed for hydrostatic bearings to provide the associated improved qualities. Taking advantage of this the motorized spindle was supported in two six-pad conical bearing bushes Fig. 10. Bearing clearance was set by adjusting the end caps to provide the appropriate axial float of the spindle. Machining of the bearing pockets was carried out in a single spark-machining operation to the layout of Fig. 10. In operation oil was fed to each pocket through a capillary tube fitted in the bearing bush. The exhaust oil returned by gravity was sealed from escape at the grinding wheel nose by a two-land labyrinth. Most of the comments on the four-pad journal bearing also apply to the six-pad conical bearing. The calculations can be carried out by a similar procedure and this is outlined in the appendix. In the following, attention is drawn to the differences from the previous section. (i) Oil flow in the unloaded condition go = nB\u00df (dimensionless) B = \u03c0 (D1 + D2) 12 nh (ii) Theoretical temperature rise Because of the conical shape an integration must be carried out to determine the energy dissipation NF resulting from the sliding speed in one bearing bush, i",
+            " ROWE The torque required to overcome friction in the bearings may then be calculated simply from Torque = \u2122* (lbf/in.) where \u03c9 is the rotational speed in rad/sec. At a speed of 7000 rev/min the torque required to turn the spindle described would be 14 lbf/in. (iii) Unloaded stiffness \u03bb\u03bf Stiffness may be evaluated for radial loading or axial loading. Considering the case of radial loading and a = 13\u00b0 \u03bb\u03bf = 4-3)8(1 -j8) l + 0 - 5 y ( l - j 8 ) and if \u03b2 = 0-5 and y = 0-5 then \u03bb0 = 0-955, which is 26% improvement over the four-pad design. The stiffness is optimized when imum \u2014 1 T i r \\y2 y) Referring to Fig. 10. The effective area should be calculated according to A'e = i (L \u2014 h) (Di + D2) cos a (iv) Load-bearing capacity ffWx The load-bearing capacity was found by the method given in the Appendix to be 2 Wm&x = 1 + ( 1 + y (1-/3)) ( 1 - 0 ) ( l + 2 - 9 6 y ( l - f l ) [l + ( \" ^ - ^ (0-25 + y)] (l + (-~^ (7-48 + 2-9&y)\\ (1 - l - 7 2 y ( l - f f l 1 + ( l - j 8 ) (8 (3 \u202251 - l-72y)) Grinding Machine Spindles 465 and if \u00df = 0-5 and y = 0-5, then J?max = 0-69, which again indicates a 26 % improvement over the four pad design",
+            " Not much of this increased accuracy is lost if the pressure in the pocket adjacent to a pocket under consideration is approximated (for the circumferential flow effect only) by the following method, P 2 - P o - ^ ( P i - P o ) . From which \u03c12 = 0\u00b74\u03a1\u03b9 + 0\u00b76\u03a1\u03bf 476 W. B. ROWE and hence Pi == Ps. Evaluating the other pocket pressures and with due respect to areas ( l + y ( l - \u00a3 ) ) (1+2-96 y ( I - / ? ) ) 2 t f W = l + 1 + ( - ~ ^ (0-25 + \u03b3\u03ae ( l + (^j^- (7-48 + 2-96 \u03b3\u03ae ( 1 - 1 - 7 2 y ( 1 - 0 ) ) 1 + ( 1 Q \u00ae (3-51 - 1 - 7 2 / ) ) (A13) \u00df I f \u00df ^ 0-5 and y = 0-5, ^max = 0-690 i.e. a 26 % improvement compared to a four-pockets bearing. (c) Conical Journal Bearing with Six Pads. Capillary compensated (Fig. 10) When a conical bearing is viewed in a diametral plane the four bearing sections should not lie tangentially to a common circle, because there will be no resistance to rotation about the circle axis. Providing this effect has been made negligible by attention to the cone angle in relation to the shaft length the analysis may be carried out simply. Gap changes calculated previously from eqn. (A5) for cylindrical bearings are calculated from the modified expressions. COS a P \u0394 Aradial loading = \u00df __ \u00df (fr \u2014 e COS 0)3 COS \u0398 dd Also C/2 \u0394 \u00c4axial loading = \u03b2 _\u03b2 \\ (n ~ e c o s \u03b8)3 \u03ac\u03b8 i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002777_cdc.2012.6426745-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002777_cdc.2012.6426745-Figure5-1.png",
+        "caption": "Fig. 5. Optimal Dubins Trajectory",
+        "texts": [
+            " However, if there is an intersection between the group trajectory and an obstacle circle, then the trajectory replanning is carried out to repair the selected segments to avoid the intersected obstacle circles. Dubins trajectory planner has been applied to generate constant speed nonholonomic kinematic feasible trajectories for a single nonholonomic vehicle [13][19][20]. Essentially, a Dubins curve is a 3-segment curve that connects between two specified poses, and each segment of a Dubins curve is either a circle arc or a straight line (see Figure 5). For every pair of specified poses, there is a set of six possible Dubins curves: {RSR,RSL,LSL,LSR,RLR,LRL}. R denotes a right turn maneuver, S denotes a straight maneuver, and L denotes a left turn maneuver. The left and right turn maneuvers are based on two turning circles located at the left and right sides of the pose (see Figure 5). C\u2212 Start and C+ Start denote the left and right turning circles of the start pose, and C\u2212 Final and C+ Final denote the left and right turn circles of the final pose. The figure depicts a LSR Dubins curve connecting from the start pose to the final pose. To determine the optimal Dubins trajectory that connects a pair of specified poses, the algorithm selects the shortest-length curve from the set of six possible Dubins curves. Readers can refer to [19] for more details about the Dubins planner that has been applied to fixed-wing UAVs",
+            " To replan the Dubins-based trajectory, we exploit segments of circle arcs and lines to repair the initial optimal trajectory since these segments can be easily concatenated using Dubins segments to form a new nonholonomic kinematic feasible trajectory. Here, the circular arc of the obstacle circle is used to provide the intermediate turning segment that allows the formation to traverse while avoiding the obstacle. This concept is illustrated with a simple example depicted in Figure 6. In the figure, the optimal Dubins trajectory (see Figure 5) intersects the obstacle circle and it can be repaired by a two-step implementation to detour the intersected obstacle circle. At Step 1, the planner determines a 2-segment Dubins curve which connects the start pose to the left tangent point of the selected obstacle circle to avoid the collision with the circle. The left of Figure 6 illustrates this step. Once a collision-free 2-segment Dubins curve is determined, the planner determines the remaining segments that connects from the left tangent point to the final pose"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002963_j.mechmachtheory.2009.11.002-Figure2-1.png",
+        "caption": "Fig. 2. The relationship between the precession angle and the rotating angle.",
+        "texts": [
+            " Its location variation always corresponds to a change of rotor balance state and can be seen as an indication whether mass unbalance is responsible for the increase of vibration at the stage of fault diagnosis [2]. This can be seen from Eq. (7). If the working condition is stable, a0 is decided by and only by aw. Obviously, the relationship between a0 and aw shown in Eq. (7) is nonlinear and is inconsistent with the isotropic assumption made by traditional methods. Suppose the alteration of aw by an increment Daw results in a corresponding change of a0 by another increment Da0, see Fig. 2. Generally, Daw \u2013 Da0. The larger the eccentricity of the synchronous precession orbit is, the greater the difference between Daw and Da0 will be. Through an angle compensation procedure, we have [6] Da0 \u00bc \u00f0a02 h\u00de \u00f0a01 h\u00de Daw \u00bc arctan\u00bd\u00f0a=b\u00de tan\u00f0a02 h\u00de arctan\u00bd\u00f0a=b\u00de tan\u00f0a01 h\u00de ; \u00f08\u00de where a and b are the magnitude of the major and minor axis, respectively. Since the isotropic assumption is no longer a prerequisite, the locating of the heavy spot then could be accurate. The correctness of this angle compensation has been proved theoretically and experimentally with successful applications [5,6]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001219_s12541-009-0102-4-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001219_s12541-009-0102-4-Figure10-1.png",
+        "caption": "Fig. 10 Virtual prototype of the ultra-precision machine",
+        "texts": [
+            " The thrust compliance of the C-axis feed system was estimated from a normal load applied to the center of the C-axis table and the average value of thrust displacements measured at the four measurement points, and the radial compliance of the C-axis feed system was evaluated from a radial load applied to the side of the C-axis table and the average value of radial displacements measured at the four measurement points. As the measurement results, the thrust and radial compliances of the C-axis feed system were revealed to be 0.0031 \u00b5m/N and 0.0224 \u00b5m/N, respectively. A virtual prototype of the ultra-precision machine for machining large-surface micro-features, which was constructed based on ANSYS software to estimate its compliances, is presented in Fig. 10. The virtual prototype was composed of 134,729 nodes, 534,911 solid elements, and 408 matrix elements. The matrix elements were introduced in order to represent the normal and lateral stiffnesses of the hydrostatic guideways, the radial and thrust stiffnesses of the hydrostatic bearings, and the axial stiffness of the linear motors. As the boundary condition for the structural analysis and measurement, the movement of the ultra-precision machine was restricted in the vertical direction at the four supporting points of the bed, as shown by the red arrows in Fig. 10. The bed, column, cross beam, and feed tables were made of cast iron (GC300), the C-axis shaft and bracket were made of steel (SCM440, SS400), and the mover/rotor and stator of motors were made of Fe-Si. Table 3 shows the material properties used for the virtual prototype. An external force inevitably causes structural deformation in the ultra-precision machine for machining large-surface microfeatures. For cases that the external force is a static load, the relationship between the stiffness matrix K of the virtual prototype, displacement vector u, and external force vector f can be expressed as Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001618_s0580-9517(08)70600-7-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001618_s0580-9517(08)70600-7-Figure1-1.png",
+        "caption": "FIG. 1. CATHODE COMPARTMENT",
+        "texts": [
+            " This, when plotted against time, would yield, upon integration of the area under the curve, the capacity of the system to do work or the coulombic output. The problem, in retrospect, of using metabolizing bacteria to obtain this capacity factor was readily solved. It was decided to utilize E. coli for the studies as it would then be possible to correlate its known metabolic behaviour with the electrochemical data. As two half cells are required to construct an electrochemical cell, the apparatus shown in Fig. 1 was developed and used in the initial studies. In a typical experiment the anode compartment of the electrochemical cell contains 20 ml of the E. coli suspension prepared as described under 111-F diluted with 20 ml of the pH 6-7 phosphate buffer. T o the cathode chamber is added 50 ml of 0.01 M potassium ferricyanide in the same buffer. The cathode serves as a nonpolarizable counter-electrode. The potential of this electrode when compared to a saturated calomel reference electrode remains essentially constant throughout the course of an experiment",
+            " coli will be described to illustrate the kind of results that may be obtained by studying microbial systems under carefully selected abnormal conditions. Chlorpromazine (CPZ) has been reported to be a general depressant of cellular metabolism (Decourt, 1955). In view of this effect, the action of CPZ on the coulokinetic behaviour of E. coli was investigated (Allen, 1966a). The coulombic outputs obtained with various concentrations of IX. CELLULAR ELECTROPHYSIOLOGY 277 * Amount (milligrams) per 45 ml of anolyte (micromolar equivalents in parentheses). CPZ when the micro-organisms metabolize glucose (macro cell, Fig. 1) are described in Table V. Although the coulombic values differed from the control in the CPZ concentration range from 0.5 to 4.0 mg, the time required for a current response and the rate of current increase (dI/dt) after addition of the glucose were essentially the same as for the control, i.e., 4 min and 210 PA/ min, respectively. With the 8.0- and 10-0-mg CPZ concentrations, 27 min elapsed before a current response was noted. The dI/dt was 70pA/min. The maximal current (I) in all cases was 760 to 780 PA",
+            " As the rate determining step in the metabolism of lactose is its transport via a permease, it is logical to assume that the initial increasing portion of the I / t curve up to the steady state metabolic rate (maximum I ) , is directly related to the transport of lactose to the site of enzymatic hydrolysis. As expected using bacteria in which p-galactosidase and permease had been induced during growth, the rate of increase of current with time using the lactose substrate was considcrably faster than that obtained with bacteria grown on glucose. Current measurements were conducted with enzyme induction in the presence of various concentrations of chloramphenicol as shown in Fig. 1 l a and correlated with rate of formation of /?-galactosidase (1 lb). 280 M. J. ALLEN It was felt that it would be advantageous to reduce the I / t curves (Fig. 1 la) to a numerical value. Since the rate of induction is directly related to the increase of dI/dt with time (d2I/dt2), a convenient measure of induction was the \u201cinduction factor\u201d. The \u201cinduction factors\u201d were the averaged A(dI/dt) values starting at zero time and continuing up to the end of the initial maximum current rise period. These \u201cinduction factors\u2019\u2019 in ,uA/min IX. CELLULAR ELECTROPHYSIOLOGY 1 28 1 700 600 500 400 300 200 100 0 500 200 100 FIG. l la . 60 120 180 240 300 360 420 480 540 600 T l h l C l n i i n u t e s l FIG. llb. FIG. 1 1. Correlation of current/time with induction of /3-galactosidase/time. Curves obtained with lactose metabolizing Escherichia coZi (glucose grown) in the presence of various concentrations of chloramphenicol. 1. control; curves 2-6 represent the effect of 0.5, 1.0, 1.5, 2.0, 2.5 mg chloramphenicol, respectively, per 25 ml anolyte. , are given in Table VII. If the averaged rate of increase of 8-galactosidase activity with time is correlated with the \u201cinduction factors\u201d, a value of 0.61 units /?-galactosidaselmin is obtained which is equivalent to an \u201cinduction factor\u201d of 1 ,uA/min"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure4-1.png",
+        "caption": "Figure 4. Arc-driving path, when l1>=l2.",
+        "texts": [
+            " Depending upon the geometry of the target direction line, the following different cases will be considered: Case 1. (Py \u2265 0 and |\u03b8| \u2264 90o) This is the case when the y-position of P, Py is positive, and the directional difference value, theta, is between the current and the target directions is less than 90 degrees. Here, again, two subcases can be considered, one subcase is when l1 >= l2 and the other is when l1 < l2, where l1 denotes the distance from point P to the current position and l2 denotes the distance from point P to the target position. Case 1-1 (Py \u2265 0 and |\u03b8| 90o and l1 >= l2) Figure 4 shows a case when l1>=l2 As indicated in Figure 4(b), the vehicle should travel from the current position, C(0,0,0), straight up to point A (equivalently, a path of infinite radius) and then follow an arc of radius R1 in a clockwise direction until the target point T is reached. The radius value, R1, can be determined by finding a circle that passes through the target point T and at the same time is tangent to the target direction line and also to the current direction line, which is the Yw axis. R1 is calculated as follows; (1) Since, PA=PT=TB/sin(\u03b8)=|a|/sin(\u03b8) and R1=CA=PA tan(\u03c8)=|a|/sin(\u03b8)\u00d7tan(\u03c8) where, (2) The y-coordinate position of point A(0, Ay) (which is on the Yw axis) at which the arc path starts is given by Ay=Py\u2212PA=Py\u2212|a|/sin(\u03b8) (3) Case 1-2 (Py \u2265 0 and |\u03b8| \u2264 90o and l1 < l2) Figure 5 shows the case of l1<l2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003073_aim.2011.6027073-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003073_aim.2011.6027073-Figure4-1.png",
+        "caption": "Fig. 4. Nonholonomic mobil e robot with two actuat ed wheels [13]",
+        "texts": [
+            " In CCSA, the hi is expressed as an antibody and (7) is expressed as an ant igen. 1 1 0 { ( ), ( )} d i i i i i L length P h P h+ + = = \u2211 (7) The simulat ion results will be presented in sect ion V. B. Trajectory of a nonholonomic mobile robot with two actuated wheels 1) Define a problem As shown in Fig. 2 the environment of prob lem is same as the previous example but mobile robot has constraints which are related to kinemat ic of robot. To th is prob lem kinematic of wheeled mobile robot shown in Fig. 4 is applied. In this figure {XOY} and {xOLy} are global and local coordinate respectively. The OL is located on the midd le of right and left driv ing wheels. The OC is the center of robot mass which has a d distance from OL on the x-axis. The configuration of mobile robot can be described by five generalized coordinates represented below [14], where (X,Y) are coordinates of OL in the global coord inat ion, \u03c6 is the heading angle of mobile robot, and \u03b8r, \u03b8l are the angles of right and left d riv ing wheels"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.2-1.png",
+        "caption": "Fig. 3.2. Representation of vectors needed for the computation of the velocity contribution of a revolute joint to the end-effector linear velocity",
+        "texts": [
+            " Each term represents the contribution of the velocity of single Joint i to the end-effector linear velocity when all the other joints are still. Therefore, by distinguishing the case of a prismatic joint (qi = di) from the case of a revolute joint (qi = \u03d1i), it is: \u2022 If Joint i is prismatic, from (3.20) it is q\u0307ijPi = d\u0307izi\u22121 and then jPi = zi\u22121. \u2022 If Joint i is revolute, observing that the contribution to the linear velocity is to be computed with reference to the origin of the end-effector frame (Fig. 3.2), it is q\u0307ijPi = \u03c9i\u22121,i \u00d7 ri\u22121,e = \u03d1\u0307izi\u22121 \u00d7 (pe \u2212 pi\u22121) and then jPi = zi\u22121 \u00d7 (pe \u2212 pi\u22121). For the contribution to the angular velocity , in view of (3.18), it is \u03c9e = \u03c9n = n\u2211 i=1 \u03c9i\u22121,i = n\u2211 i=1 jOiq\u0307i, (3.28) where (3.19) and (3.23) have been utilized to characterize the terms q\u0307ijOi, and thus in detail: \u2022 If Joint i is prismatic, from (3.19) it is q\u0307ijOi = 0 and then jOi = 0. \u2022 If Joint i is revolute, from (3.23) it is q\u0307ijOi = \u03d1\u0307izi\u22121 and then jOi = zi\u22121. In summary, the Jacobian in (3.5) can be partitioned into the (3 \u00d7 1) column vectors jPi and jOi as J = \u23a1 \u23a3 jP1 jPn "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001141_j.jmatprotec.2008.02.071-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001141_j.jmatprotec.2008.02.071-Figure13-1.png",
+        "caption": "Fig. 13 \u2013 Rolling position of the largest pressure angle for the rack cutter.",
+        "texts": [
+            " If the polar angle of point B is solved, the x-component of point B is Bx = rB cos and the y-component is By = rB sin . From the geometric relationship shown in Fig. 12, can be expressed as = \u2220KOE + \u2220EOB (56) where \u2220EOB = inv B \u2212 inv E (57) Due to pure rolling between the pitch line of the rack cutter j o u r n a l o f m a t e r i a l s p r o c e s s i n g t e c s \u0131 w t F t t d 5 A t c t a c r onstant outside diameter, and variable outside diameter ccording to the shifted amount, respectively. hown in Fig. 13. The angle \u0131=\u2220NOJ is therefore = r1 (58) here r1 is the radius of the pitch circle of the generated spuryped cutter. Therefore, angle is obtained as = inv B \u2212 inv J + \u0131 (59) ig. 14 shows the width of the top land of the spur-typed cuter when the outside diameters are kept constant, and when he outside diameters are changed according to the shifted istance, respectively. . Conclusion novel design for a hob cutter capable of generating a spuryped cutter in one hobbing process has been proposed. The utting edge in the normal direction has been designed as hree straight lines with different pressure angles and two rcs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001260_epepemc.2008.4635471-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001260_epepemc.2008.4635471-Figure1-1.png",
+        "caption": "Figure 1. Laboratory magnetic levitation system",
+        "texts": [
+            "ndex Terms\u2014Estimation technique, non-linear control, magnetic device, real time processing, motion control. We consider a laboratory magnetic levitation system presented in the figure 1. This plant consists of an electromagnet, an optic position sensor, a ferromagnetic sphere and a power interface. The power interface is used for operation of electromagnet coil, and subsequently the generation of a magnetic field that will lift the ferromagnetic sphere and allow its levitation. The measuring unit consists of the strong visible light source and the optical sensor with an appropriate electronic circuit. The sphere covers the sensor which measures the degree of light dousing. Because of the spherical shape the measurement is nonlinear and needs to be electronically or digitally linearised. This system was considered among the others in the following works [1], [5]\u2013[7], [9], [15]\u2013[19], [21], [22]. The ball velocity measurement is a difficult task, because the only possible options is a numerical differentiation of the position. That is why an observer design is such an attractive proposition. Among the others the observers for magnetic levitation system were considered in [3], [4], [20]. Magnetic levitation system depicted in the figure 1 can be described with the following differential equations x\u03071 = x2 x\u03072 = 1 2m dL(x1) dx1 x2 3 + g (1) x\u03073 = 1 Ts (ksu \u2212 is \u2212 x3) 1Work financed by state science funds for 2008-2011 as a research project. Contract no. N N514 414034 where: x1 - ball position [m], x2 - ball velocity [m s ], x3 - current [A], u - control voltage [V ], m - ball mass [kg], L(x1) - electromagnet inductance [H ], g - gravitational acceleration [ m s2 ], ks - gain of current controller [ A V ], Ts - time constant of the linear current driver [s], is - zero error of current controller [A], For the purpose of this paper we use the following approximation of the electromagnet inductance (see [18]): L(x) = L0e \u2212 x1 b (2) As it can be seen in the figure 1 the ball position is measured by the appropriate sensor and the coil current also is easily measured. The velocity measurement, however, is difficult to realise. That is why in this paper we consider velocity estimation with the use of only the measured position, and with the use of both measurements of position and current. However at first we will want to determine system observability. We will verify, is the system observable (see also [3], when only the position is available (because if it is, then additional use of current measurement will be used as improvement)",
+            " The main idea of our approach is to realise (14) as a nonlinear cascade with a separate controller for tracking of coil current signal (x3) that it would be appropriate for feedback linearisation. This control structure (supplemented with reference value and the observer is presented in the figure 5. 1448 2008 13th International Power Electronics and Motion Control Conference (EPE-PEMC 2008) All of the considered control structures were validated experimentally. The laboratory system owned by the Department of Automatics of AGH, depicted on fig. 1 was used for all the experiments. It\u2019s measurements and the actuator were connected via the PCI converter card to a Windows based PC with MATLAB/Simulink environment. Experiments were performed in Simulink with RTWT. Numerical derivatives for position PID controller were realised with forward differences. Observer equations were solved with a 5th order Runge-Kutta formula implemented in Simulink. Discretisation step was 700\u03bcs for all of the experiments. Values of position on following plots should be interpreted as a distance from the electromagnet, so the lower the value the higher ball was"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001455_6.2008-6462-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001455_6.2008-6462-Figure3-1.png",
+        "caption": "Figure 3. Local navigation system",
+        "texts": [
+            " 3 of 14 American Institute of Aeronautics and Astronautics This article is organized as follows: first, the guidance system is presented, second, the autopilot design using CIHMPC is discussed including theoretical considerations, finally, simulation results are presented and discussed. The objective of the guidance system is to calculate lateral and vertical deviations from the actual position of the aircraft to a navigation leg based on two consecutive waypoints and, based on these informations, provide lateral and vertical guidances to the autopilot. The system shall also be able to automatically change the active leg based on proximity to the next waypoint in order to smooth the transition. The navigation legs are straight lines between two consecutive waypoints. Figure 3 shows the geometry of the proposed navigation. The aircraft\u2019s present position is called PP; the past waypoint overtaken by the vehicle is called PW and the active waypoint ahead is called AW. Each leg has a local reference system northeast-down (NED) centered on PW where all other positions are referenced to. Initially the guidance system reads PP (as vehicle output), PW, and AW in WGS-84 geodetic coordinates latitude lat, longitude lon and altitude H and performs a transformation to ECEF (earth-centered earth-fixed) cartesian coordinates according to Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001491_robot.2008.4543713-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001491_robot.2008.4543713-Figure3-1.png",
+        "caption": "Fig. 3. (a) Asymmetric lens and (b) quasioptimal flower coverage",
+        "texts": [
+            " When sensing circles ci is intersecting with sensing circles cj, k21j ,,, L= , the summation of the area of k asymmetric lenses is literally the coverage overlaps. We adopt the same notation \u03b5j as the central angle of the segment of ci covered by cj for k21j ,,, L= . According to Lemma 6, the area of intersection is strictly decreasing in rj for fixed ri and central angle \u03b5j. In order to reduce these lens areas, and thus, the total coverage overlaps, we define the central angles by \u03c0\u03b5 2 r r k 1i i j j \u22c5= \u2211 = (4) which means the central angle \u03b5j is proportional to the neighbor radius ri. Obviously, (4) implies \u2211 = = k 1j j 2\u03c0\u03b5 . As shown in Fig. 3(b), when all the interstices are just vanished, the summation of central angles equals 2\u03c0. Therefore, (4) suggests a configuration of flower to achieve quasioptimal flower coverage. Moreover, the central angles \u03b5j are assigned in proportion to petal radius rj, the area of overlaps is reduced due to Lemma 6. The desired distance d\u0302 is: 2 r 2 rd j j j iij ' coscos\u02c6 \u03b5\u03b5 += (5) where \u239f \u239f \u23a0 \u239e \u239c \u239c \u239d \u239b = \u2212 2r r 2 j j i1 j \u03b5 \u03b5 sinsin' . (6) If vi is a boundary node, the desired distance with its neighbor vj are simply defined as )(\u02c6 jiij rr 2 3d += "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003064_978-3-642-34336-0_18-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003064_978-3-642-34336-0_18-Figure2-1.png",
+        "caption": "Fig. 2. Scheme and axis movements of a bevel gear cutting machine",
+        "texts": [
+            " These challenges and issues are presented in figure 1. In order to provide a simulation tool for analyzing and optimizing the bevel gear cutting process a manufacturing simulation has been developed at WZL, as presented by [2-3]. Within the manufacturing simulation a geometrical penetration calculation is conducted as described by [2-4]. Before starting the calculation the kinematics , workpiece and tool have to be modeled as described below. All axis movements, like in the real cutting process, can be considered in the simulation. In figure 2 the axis movements and positions of the bevel gear cutting machine are presented. More precisely, the variables and the resulting movements are: mccp (the machine center to cross axis point), \u03b1 (cradle angle), \u03b2 (modified roll by workpiece rotation), \u03b3 (angular Motion), \u03c6 (radial motion), \u03b5 (horizontal Motion), \u03b7 (vertical motion), \u03c7 (helical motion), \u03c9 (tool rotation). All axis movements can be described as a series expansion, which is truncated after the sixth order, regarding the mean tool angle \u03c9m and the mean cradle angle \u03b1 m"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002348_1350650111427508-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002348_1350650111427508-Figure3-1.png",
+        "caption": "Figure 3. Layout of the double Lomakin hybrid bearing with counter-rotating injection.",
+        "texts": [
+            " The floating bearing design is based on the assumption that the tested bearing is free to move in X and Y directions (Figure 2), while the shaft can only rotate around its axis. The hybrid bearings supporting the shaft must then have stiffness orders of magnitudes larger than that of the tested bearing; otherwise, its displacement will interfere with the vibration of the bearing housing. The shaft is therefore guided by two identical externally pressurized bearings of Lomakin type depicted in Figure 3. Four orifices inclined at 85 against shaft rotation inject water pressurized up to 75 bar into a circumferential groove located in the bearing midplane. Due to the high feeding pressure, each waterlubricated Lomakin hybrid bearing is characterized by high radial stiffness. In addition, the rotor has a large tilting stiffness because the two Lomakin bearings are largely spaced. The angled injection ensures a very low or negative cross-coupling stiffness, leading to a large effective damping. Hence, the test rig was operated up to 60 kr/min (the largest speed enabled by the spindle) without any sign of instability"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002977_robot.2010.5509531-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002977_robot.2010.5509531-Figure1-1.png",
+        "caption": "Fig. 1. Spoke-like passive-dynamic walker that consists of two identical crossed frames",
+        "texts": [
+            " In limit cycle walking, the stance leg is instantaneously exchanged if the heel strike collision model is inelastic. In dynamic walking of humans and biped humanoids controlled based on zero moment point (ZMP), double-support phase plays an important role in stabilization of the gait generation. Also in limit cycle walking, the effect of double-support phase control must be effective to improve its stability and robustness. The author has developed a novel spoke-like passivedynamic walking machine shown in Fig. 1 to investigate the effects of double-support phase and swing-leg retraction [4][5][6][7][8] on the gait efficiency and stability. This machine appears something like an rimless wheel [9], but is a dynamic bipedal walker whose swinging motion is emerged by the viscosity. By taking the special configuration, we can conduct experimental study very easily without concerning the foot scuffing at mid-stance. We have a prediction that this walker would exhibit passive-dynamic walking incorporating double-support phase and exchange the stance leg more smoothly if the leg frame has elastic elements",
+            " Another goal of this study is to develop an efficient and robust locomotion system that can walk on uneven surface and can climb up a steep slope at high speed. Spoke-like locomotion machines are expected to achieve high adaptability to complex outside-environment [10]. Although this paper does not report, we are now considering to add a telescopic actuation reported in [11]. Although we have not investigated the passive-dynamic gait with double-support phases yet, several interesting results have been obtained through inelastic collision model corresponding to the walker shown in Fig. 1. This paper then reports the results; the mathematical modeling, the simulation results and the experiment of passive-dynamic walking on a treadmill. The dynamic walker does not have any actuators and consists of two identical crossed frames as shown in Fig. 1. The blue leg frames are attached to the outer axis-frame, whereas the aqua leg frames are attached to the inner axisframe. These two wheels can rotate independently around 978-1-4244-5040-4/10/$26.00 \u00a92010 IEEE 1703 the central axis, however, the viscous friction between them can be changed by the adjust screw shown in Fig. 2 left. End of each leg has a rubber for shock absorbing. Although we do not describe the detail, this walker is also able to run by replacing the leg frames with telescopic leg frames incorporating a spring as shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000815_mcs.2008.927961-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000815_mcs.2008.927961-Figure3-1.png",
+        "caption": "FIGURE 3 Rotational and translational motion. (a) Rotational motion. Each wheel is oriented 45\u25e6 or \u221245\u25e6 with respect to the vehicle body, and all four wheels spin at the same rate. The vehicle body thus rotates around its geometric center at a fixed rate with zero turning radius. (b) Translational motion. Each wheel is oriented at the same angle with respect to vehicle body and the global reference frame, and all four wheels spin at the same rate. Thus, the vehicle body translates in the global reference frame without a change of attitude.",
+        "texts": [
+            " The model can be reduced to an inertia link in cases of small armature inductance. (b) 1 \u03a9 (s) \u2212 \u2212 + + Td (s) Tm (s)Ia (s)Ua (s) Kt Ke Js + B 1 Jas + Ra (a) 22 IEEE CONTROL SYSTEMS MAGAZINE \u00bb OCTOBER 2008 four wheels that are individually driven and steered, thereby making it omnidirectional. Unlike other agricultural navigation platforms, the vehicle can perform motions that are either purely translational or purely rotational. We first analyze the translational velocity and rotational velocity vector for each wheel as shown in Figure 3 and then study the composition of these velocity vectors as shown in Figure 4. During purely translational motion, the longitudinal directions of all four wheels are oriented identically with respect to the vehicle body, and all four wheels spin at the same rate around their drive axes. Therefore, as shown in Figure 3(b), all of the wheel centers, and thus the geometric center of the vehicle body, move at the same velocity. For purely rotational movement, the longitudinal axis of each wheel is oriented at either 45\u25e6 or \u221245\u25e6with respect to the orientation of the vehicle body, as shown in Figure 3(a). Due to the geometry of the vehicle, the velocity orientation of each wheel center is perpendicular to the line between the vehicle geometric center and the respective wheel center. Since all of the wheels are commanded to rotate at the same angular velocity, the vehicle rotates as a rigid body around its geometric center. Consequently, the vehicle\u2019s turning radius is zero, giving the vehicle maneuverability that is superior to most two-wheel-steered agricultural vehicles. We now combine the individual wheel translational and rotational motions to obtain the orientation and velocity of the vehicle as a whole"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003836_1.3552507-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003836_1.3552507-Figure4-1.png",
+        "caption": "FIGURE 4: hybrid analysis of strain for gear rolling process",
+        "texts": [
+            " Thereby the defined deformation during rolling process, specifically the phased penetration of the tools into the rotationally symmetric preform was analyzed. A range of penetration rates (1.5 mm, 2.5 mm, 3.5 mm, 4.5 mm) and two different work piece materials (C15, 16MnCr5) were performed. The penetration speed was set to 0.05 mm/s and the rotational speed of the tool to 6 rpm (see figure 3). To use these data as efficiently as possible for subsequent studies a hybrid approach to this problem was chosen. In that case the combination of visioplastic determination and FEM simulation (see figure 4). The deformation analysis is using multiple cameras as videometric system (in this case four), which reproduce the examined measurement grid. The automatic image analysis follows the electrochemically applied grid lines (original grid spacing x = 2 mm) and can identify identical lines in all images when assigned interactively at the beginning of a common starting point. From the image coordinates of the crossing points the spatial coordinates of the entire deformed measurement grid can be determined, exactly when the fixed camera configuration is measured in advance or calibrated"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003710_gt2012-68956-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003710_gt2012-68956-Figure3-1.png",
+        "caption": "FIGURE 3. Overview over the test rig",
+        "texts": [
+            " In the 1990\u00b4s, a test rig was erected at the Institute of Jet Propulsion and Turbomachinery of the Unversity of Braunschweig, Germany. This test rig is used for investigations of brush seals under realistic steam turbine conditions. However, it was found that the behavior of brush seals under steam cannot be understood by evaluating the results of this test rig only. Therefore, a second rig was erected which allows testing individual brush seals by means of quality characteristics. This test rig is depicted in FIGURE 3. 2 Copyright c\u00a9 2012 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/74826/ on 01/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Depending on the brush seal type, a maximum of three brushes can be tested within one setup. The seal carrier itself can be rotated around the rig axis which allows observing brush characteristics at reproducible circumferential positions. However, a rotation of the seal carrier during testing sessions is not possible"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure2.9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure2.9-1.png",
+        "caption": "Figure 2.9 Hydraulic actuator block diagram",
+        "texts": [
+            " (Intuitively this makes sense. Consider steering your car. As you oscillate the steering wheel angle very slowly from left to right the car will make large oscillations in heading. But when you waggle the steering wheel quickly over the same amplitude the car will wobble a little but the basic heading will not change significantly.) To illustrate these principles using control terminology, let us consider an example based on the control valve integration referred to above. Referring to the block diagram of Figure 2.9, we have a control valve (shown as a gate valve with a linear position xV) controlling the flow 34 Closing the Loop into a half-area actuator. Supply pressure PS is applied to the control valve and also to the half-area side of the servo-actuator piston. A fixed orifice connects the control pressure side of the control valve PC to the hydraulic return pressure PR, which we can assume to be nominally zero. This arrangement allows movement of the control valve to cause movement of the piston in either a positive or negative direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.6-1.png",
+        "caption": "Fig. 1.6. Effective front axle characteristic Fyl(aal ) influenced by steering compliance.",
+        "texts": [
+            " The front tyre slip angle is denoted with a~. The 'virtual' slip angle of the axle is denoted with aal and equals (cf. Fig. 1.5): O~al - 0~1 - ~ffcl (1.24) where both a~ and ~c~ are related with Fy~ and Mz~. The subscipt 1 refers to the front axle and thus to the pair of tyres. Consequently, Fy~ and Mz~ denote the sum of the left and right tyre side forces and moments. The objective is, to find the TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 13 function Fyl(aal ) which is the effective front axle characteristic. Figure 1.6 shows a graphical approach. According to Eq.(1.24) the points on the Fy~(al) curve must be shifted horizontally over a length ~cl to obtain the sought Fyl(a~l). The slope of the curve at the origin corresponds to the effective axle cornering stiffness found in the preceding subsection. Although the changes with respect to the original characteristic may be small, they can still be of considerable importance since it is the difference of slip angles front and rear which largely determines the vehicle's handling behaviour"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003360_gt2010-23802-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003360_gt2010-23802-Figure1-1.png",
+        "caption": "Figure 1 Schematic view of component level test rig including detailed view of bearing housing.",
+        "texts": [
+            " Importantly enough, the curvefit of the simulated data with the KCM model rendered added mass coefficients of about 2000 kg, approximately 30 times the mass of the rotor assembly. On the other hand, the largest added mass coefficient from Ref. [8-12] identified from actual experimental data is ~70 kg (i.e. the same order of magnitude as the test device). The present experimental work aims to identify the dynamic force coefficients for two sets of rocker pivot TPJBs while assessing frequency dependency of the force coefficients. The following section describes the experimental facility used to evaluate oil lubricated journal bearings. This component level test rig (Fig. 1) was developed to measure rotordynamic force coefficients on oil-lubricated tilting pad bearings and is described in Ref. [18]. Although engineered with some differences, the general test rig layout and testing concept draws upon past work conducted by Childs [19]. In practice, a tilting pad bearing is usually rigidly mounted into the machine casing and experiences forces generated from the rotor weight and rotor imbalance. Contrary to this typical operating condition, the test rig shown in Figure 1 flexibly mounts the test bearing, where the primary forces into the bearing are precisely applied externally through the housing and not internally from the rotor. The precision balanced test rotor (Fig. 1) was driven by a 20krpm 100kw AC motor and was straddle mounted by using two sets of angular contact ball bearings (slave bearings), which provided a hard mounted short rotor configuration. This fact enabled the test rotor and slave bearing system to have first critical speed that resided at 32krpm, well beyond the test frequency range of the test bearing. This critical speed was predicted and verified via pin-pin impact tests. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess",
+            " There are three pitch stabilizers on each axial face of the test bearing housing and serve the following functions: 1) to react against the torque developed from the test bearing, 2) to align the center axis of the test bearing with the central axis of the test rotor to provide radial compliance to the system ensuring translational movement of the test bearing with respect to the test rotor, and 4) to axially restrain the test bearing. The forced excitation into the system was delivered through the bearing housing using two orthogonally mounted electro-hydraulic shakers (exciters), which enabled the precise control of the excitation frequency and amplitude. The static operating position of the test bearing with respect to the test rotor was established with the static loader. As shown in Fig. 1 the static loader uses a hydraulic piston in combination with a pulley system in line with a soft spring pack. The electro-hydraulic exciters were used to super impose dynamic force on the static force (from the static loader), where the peak spectral amplitude for each frequency component was targeted at 2.5 microns. The test bearings are shown in Fig. 2 and consist of four different configurations of interest. Both 4 and 5 pad rocker pivot tilting pad bearing were tested where the 4 pad was tested in LBP configuration and the five pad bearing was tested in the LOP configuration. For each of these two bearing configurations, both 50% pad offset and 60% pad offset, were investigated with differing pad preload coefficients (shown in Fig. 2). Oil was delivered through the bearing housing oil-inlet port (Fig. 1) and then through oil nozzles positioned between tilting pads (Fig. 2). Several measurements were taken to both define an operating condition and also to calculate rotordynamic force coefficients. Operating conditions were established using the static load, oil inlet temperatures, rotor speed, and oil flow rate. Another measurement of interest was pad temperatures. For each bearing configuration three pad temperatures were measured near the trailing edge of the pad, offset from the edge by 12.25o. The calculation of rotordynamic coefficients required dynamic measurements (Fig. 1) of housing motion with respect to the test rotor, dynamic forces on the bearing housing, and Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/gt2010/70410/ on 04/03/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use 5 Copyright \u00a9 2010 by ASME also the housing acceleration. Motion probes were installed at the FWD and AFT end of the bearing housing, which were used to ensure a translational mode shape of the test bearing with respect to the rotor during tests"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000591_physreve.75.011706-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000591_physreve.75.011706-Figure1-1.png",
+        "caption": "FIG. 1. Molecular structures of the one-dimensional circular C , linear R=SmCP and elliptic EL=B2 phases. All the molecules lying in one plane normal to Oz have the same orientation. One bent-core molecule is represented with one vector and one two-end arrow. The mean polarization P r is parallel to the molecular polarization shown in the figures by a one-end arrow. The tilt vector A r is parallel to the rotation axis turning the molecular plane from its reference orientation parallel to the vector wave k and the polarization to its actual configuration shown in the figure.",
+        "texts": [
+            " e x and e y are unit vectors parallel to Ox and Oy, respectively. The vector wave k is a phenomenological parameter that must be determined experimentally or calcu- lated within microscopic models. P r is coupled with an axial wave A r representing the tilt of the molecular plane with respect to the direction of the wave vector which constitutes a \u201cpseudoproper order parameter 17 .\u201d The molecular orientations are determined by the equilibrium amplitudes and phases of these two waves. Three \u201cunprimed\u201d phases displayed in Fig. 1 can first be stabilized when the phases of the two waves are locked: i The \u201ccircular\u201d C phase, in which P r and A r are circularly polarized. Its symmetry group 122 continuous group of a regular helix yields a structure remaining homogeneous no density wave and macroscopically isotropic in the plane normal to k . We have proposed 11 that this chiral helielectric nonsmectic phase corresponds to the 1D approximation of B7 when one neglects the long wavelength modulation normal to the main periodicity direction",
+            " PC and AC denote their real wave amplitudes, vanishing as T\u2212Tc 1/2 at the C\u2192 Isotropic transition temperature Tc. ii In the \u201clinear\u201d R phase, the right- and left-handed circular waves in Eq. 2 have the same amplitude so that; P R r = PR cos kz 0 , A R r = AR 0 sin kz . 4 Its discrete orthorhombic group Pmma describes an achiral smectic phase, which can be either antiferroelectric or anticlinic. In the former case, the bent-core molecules align with their polarizations normal to Oz, forming the antiferroelectric two-layer sequence displayed in Fig. 1. iii In the \u201celliptic\u201d EL phase, the two circular waves forming P r in Eq. 2 have independent amplitudes, yielding a chiral, antiferroelectric, and anticlinic structure with symmetry P2221. In this phase, the polarization and tilt fields, P EL r = Px cos kz Py sin kz , A EL r = Ax cos kz Ay sin kz , 5 coincide with the structure of the phase B2 5,11 . When the phases of A r and P r are no longer locked together, additional 1D \u201cprimed\u201d phases, denoted by C , R , R , and EL with symmetries 1, P21/m, Pmc21, and P21, respectively, condense spontaneously",
+            " In the smectic stable phases, the period of the density wave is half that of the vector waves. In 2D and 3D structures, density waves are induced in the same directions as the vector waves are. The molecules are located at symmetric positions in the unit cell where the density function is maximum. For instance, in the R phase, the density maxima can coincide ei- ther with those of P or with those of A . These possibilities define two configurations of the same phase, exhibiting different physical properties 11 . In the first case, the structure is antiferroelectric and not tilted Fig. 1 , whereas in the other case the structure is anticlinic with approximately zero local polarization. The same diversity arises in the higher dimensional mesophases. In a given stable phase, the symmetry analysis alone cannot foresee which configuration will be actually stable, since it depends on the sign of the coupling coefficients between polarization and density waves. Therefore, the thermodynamic analysis assigns different regions of the phase diagram to each variant of a single phase. III. SUBPHASES OF SmCP When additional vector waves with wave vectors K normal to k condense, the linear R phase i",
+            " We shall discuss its relation to the experimental situation in Sec. VI. The present section is devoted to the enumeration and description of the ordered phases arising below R. The procedure used to obtain space groups and molecular arrangements is worked out in detail in the case of the phase denoted by A1 x, which coincides with the B1 bentcore columnar phase. In R three nonequivalent directions in the x-y plane normal to k must be considered. 1 K =Kxe x is parallel to the plane of polarization in R see Fig. 1 . The star of K set of wave vectors obtained from K after applying the point symmetries of R has two branches One sees that in C and EL A r is parallel to P r and cancels in R at least where the molecules are represented . In R and EL, only the molecules located at the maxima of the density wave i.e., in the smectic planes are represented. We have pictured the antiferroelectric configuration of R, in which the molecules lying in the smectic planes are not tilted whereas molecules located between the planes are tilted ",
+            " Continuous and broken lines indicate second- and first-order transitions, respectively. a When the free energy Eq. 7 is truncated up to the fourth degree. b When the free energy is truncated up to the sixth degree. The axes represent phenomenological coefficients defined in Eq. 7 , which are combinations of two external parameters such as temperature and concentration. Varying only the temperature amounts to a move along straight lines in the diagrams. 011706-6 structure 1 is purely anticlinic and cannot be obtained from the antiferroelectric variant of R presented in Fig. 1 and observed in the SmCP phase. In contrast, 1 can appear from the anticlinic variant of R after the condensation of the density wave cos x , which thus represents the primary order parameter of the transition. The vector wave sin x e x is then an induced secondary pseudoproper order parameter. On the contrary, in case 2 , the antiferroelectric structure results from the condensation of the vector wave in the antiferroelectric variant of R displayed in Fig. 1. The molecular positions remain thus close to the maxima x=z= \u00b11/4 of the polarization, even though they can vary slightly around these maxima in the direction Ox. Therefore, the density waves are weak secondary pseudoproper or improper order parameters. In case 3 , the structure is simultaneously anticlinic and antiferroelectric, and both waves participate to the ordering mechanism on the same footing. A1 x belongs to the same symmetry class Pmma as the parent R phase does, but it has a 2D lattice instead of a 1D smectic structure",
+            " The free energy depends on two invariant polynomials, I1= 1 1 *+ 2 2 * and I2= 1 1 * 2 2 *. In contrast with the A1 x. b Waves and molecular structures in one unit cell of the phases A1 x, A2 x, B1 x, and B2 x, see Table I . The wave components normal to Oy are represented by in-plane arrows, whereas the parallel component is represented by circles. First column: Tilt fields A x ,z . Sec- ond column: Polarization fields P x ,z . Third column: Molecular configurations. The representation of single molecules is explained in Fig. 1. 011706-7 waves and traveling along Ox or Oy, studied in sections A and B there is no bilinear coupling between 1 and 2 so that a stable low-symmetry phase where 1 or equivalently in an equivalent domain 2 cancels can be stabilized. The biquadratic coupling term I2 permits the stabilization of this phase and leads, in the general case, to distinct values for the equilibrium amplitudes of the two waves. Minimizing the corresponding free energy F I1 , I2 yields three ordered states for each wave type R\u00b1, i One columnar phase for 2=0 or 1=0 in another domain denoted by I\u00b1 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-99-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-99-1.png",
+        "caption": "Figure 3-99: Angular misalignment or dirt trapped behind carrier ring caused damage to pads.",
+        "texts": [
+            " In Figure 3-97, observe the back portion of a tin-aluminum bearing showing a flat spot on the steel shell. This was caused by debris entrapped between the shell and housing. Figure 3-98 166 Chapter 3 Figure 3-97: Flat spot on steel shell caused by careless assembly. Figure 3-98: Bearing-bore deformation resulting from the careless assembly in Figure 3-97. Machinery Component Failure Analysis 167 depicts the bore of the bearing shown in Figure 3-97. Entrapped debris has deformed the bearing bore, resulting in flexure and fatigue damage. In Figure 3-99, angular misalignment of shaft or housing or dirt trapped behind the carrier ring caused damage to the pads (however, on one side only). Whenever faulty assembly is suspected or confirmed, investigate and if necessary correct the interference conditions and housing design. Correct the misalignment, discard the bearings, and fit new bearings or pads. Inadequate lubrication may be due to insufficient oil-pump capacity, inadequate header or oil passage diameters, incorrect grooving design, or accidental interruption of the oil supply"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002575_j.triboint.2011.10.005-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002575_j.triboint.2011.10.005-Figure2-1.png",
+        "caption": "Fig. 2. General contact geometry.",
+        "texts": [
+            " Leeds-Lyon Symposium on : Economic, Environmental, In an experimental examination the influence of different surface velocity directions of two contacting partners on lubricant film thickness in an EHL point contact is investigated. EHL film thickness is measured at different sum- and surface velocities, different Hertzian stresses, lubricant inlet temperatures and skew angles in a twin disc machine. Initial tests with circular contacts are performed in order to eliminate the influence of skew angle on contact geometry. The measurement of the EHL film thickness is done by a capacitance measuring method followed by an iterative analysis. Fig. 2 shows a schematic of an elliptical point contact with the semiaxes a and b. It is formed by contacting cylinders whose rotational axes are crossed with the skew angle d. The surface velocities of the contacting partners are indicated by v , 1 and v , 2. Sum velocity v , S is defined as v , S \u00bc v , 1\u00fev , 2 \u00f01\u00de Sliding velocity v , G is defined as v , G \u00bc v , 1 v , 2 \u00f02\u00de A difference in surface velocities causes a slip ratio s. For the skew angle d\u00bc01 s is calculated according to the following equation with v1 and v2 as the velocities of discs 1 and 2, respectively, and v1Zv2: s\u00bc v1 v2 v1 100% \u00f03\u00de Nomenclature a (mm) semi-major axis of contact ellipse b (mm) semi-minor axis of contact ellipse b (mm) width of test disc bH (mm) Hertzian contact width C (pF) capacitance da (mm) outer diameter of test disc f (Hz) frequency of resonant circuit FN (N) normal force FR (N) frictional force h (mm) EHL film thickness h0 (mm) film thickness in parallel gap hgem (mm) measured EHL film thickness hmean (mm) simulated mean film thickness hmin (mm) minimum film thickness L (H) inductivity pH (N/mm2) Hertzian stress R0 (mm) equal radius of curvature Ra (mm) arithmetic mean roughness s (-) slip ratio v1 (m/s) surface velocity of disc 1 v2 (m/s) surface velocity of disc 2 vG (m/s) sliding speed vS (m/s) sum velocity d (1) skew angle e0 (F/m) dielectric constant eoil (-) dielectric coefficient of lubricant exG (1) angle of sliding speed vector exS (1) angle of sum velocity vector WM (1C) bulk temperature WM,o (1C) bulk temperature of measuring disc WM,u (1C) bulk temperature of test disc Woil (1C) oil inlet temperature n (mm2/s) kinematic viscosity n40 (mm2/s) kinematic viscosity at 40 1C n100 (mm2/s) kinematic viscosity at 100 1C r15 (kg/m3) density at 15 1C EHL film thickness h is measured in a FZG twin disc machine",
+            " As a first approach the influence of slip on film thickness can be reduced to the influence of lubricant viscosity at bulk temperature at the inlet zone of the contact. Fig. 13(a\u2013c) shows an increasing lubricant film thickness with an increasing skew angle for a given bulk temperature. The absolute value of the sum velocity vector is kept constant for all skew angles. For a skew angle da01 the surface velocities 9v , 19 and 9v , 29 are calculated according to the following equations as a function of d and exS (cp. Fig. 2) 9v , 19\u00bc 9v , S9 cosexS sind \u00f06\u00de 9v , 29\u00bc 9v , S9 sinexS 9v , S9 cosd cosexS sind \u00f07\u00de For all test conditions the absolute values of the surface velocities 9v , 19 and 9v , 29 are kept constant for all skew angles. Thus the sum velocity vector v , S represents the bisecting line of the angle between the surface velocity vectors. In order to keep the absolute value of the sum velocity vector constant the absolute values of the surface velocity vectors increase with increasing skew angle. Fig. 14 shows the relation of the absolute values of the surface velocity vectors v , 1 and v , 2 and the sliding velocity v , G, which appears for skew angles da01, as a function of the skew angle d for an exemplarily constant sum velocity 9v , S9\u00bc20 m/s"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002704_robot.2010.5509581-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002704_robot.2010.5509581-Figure2-1.png",
+        "caption": "Fig. 2. Simulations for the target curve, given by the intersection of the two surfaces in Figure 1. The circles represent several initial robot configurations.",
+        "texts": [
+            " Unless mentioned otherwise, all the vectors considered in the paper are column vectors, q = [x y z]T , and \u2207 is the gradient with respect to the coordinates x, y, z. The proposed methodology is based on the existence of two functions \u03b11 and \u03b12, such that the desired curve is obtained by the intersection of the zero level set of such functions. Let D(t) be the set of points q such that at time t the functions \u03b11 and \u03b12 vanish. Thus, under certain constraints imposed on the functions \u03b1i, the set D(t) is a one-dimensional set embedded in R 3. Figure 1 illustrates this idea. Figure 2 presents ideal simulation results which illustrates our objective: a robot must converge to and circulate along the target set D(t) from every initial condition. We assume the robot kinematic model in (1). We will first assume that the target set D(t) is static. In order to achieve convergence, we use a function V (\u03b11, \u03b12): a negative definite function with continuous partial derivatives, such that its gradient is null only at the origin. The control law for convergence is given by: u = G\u2207V , (2) where G is a non null positive function, and the vector G\u2207V points toward the set of points such that \u2207V = 0, which include the set D(t), where \u03b11 = 0 and \u03b12 = 0, since V is a negative definite differentiable function"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002014_cefc.2010.5481434-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002014_cefc.2010.5481434-Figure2-1.png",
+        "caption": "Fig. 2. The magnetic field distributions.",
+        "texts": [
+            " Recently, a new class of brushless motors with PMs located in the stator, termed as flux-switching PM (FSPM) motor, has been proposed [3]. The purpose of this paper is to propose a new modular FSPM (M-FSPM) motor, which can not only retain the advantages of the FSPM motor, but also solve its problems. A new 3-phase M-FSPM machine is proposed, as shown in Fig. 1. The key is the introduction of FT teeth between the adjacent stator poles. Thus, the individual windings are essentially isolated among phases, leading to significantly enhance the FT capability. Fig. 2 shows the no-load magnetic field distributions. It depicts that the M-FSPM motor can offer the nature of phase decoupling, which is essential for FT operation. By using finite element method, the proposed M-FSPM motor has been quantitatively compared with the conventional FSPM motor. An experimental M-FSPM motor is designed and prototyped for verification, as shown in Fig. 3. This work was supported in part by grants (Project No. 60974060, 50907031 and 50807007) from the National Natural Science Foundation of China and a grant (Project No"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001311_j.matdes.2006.11.003-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001311_j.matdes.2006.11.003-Figure1-1.png",
+        "caption": "Fig. 1. Decomposition of the deformation gradient tensor.",
+        "texts": [
+            " Here, Cp is decomposed into a symmetric tensor denoted by Kp and an orthogonal rotation tensor labeled as Qp, as follows: Cp \u00bc KpQp \u00f021\u00de The plastic stretch tensor, Kp, is obtained from Eqs. (20) and (21) in terms of V and Ve, as K2 p \u00bc CpC T p \u00bc V 1 e V2V 1 e \u00f022\u00de Using Eq. (21), Eq. (20) can be written as V \u00bc VeKpQp \u00f023\u00de Eq. (23) represents the decomposition of the left stretch tensor into elastic and plastic stretch tensors and a rotation tensor. This decomposition is used through out this paper. Multiplying both sides of Eq. (23) by R and using Eq. (3), we obtain: F \u00bc VeKpQpR \u00f024\u00de Eq. (24) is demonstrated graphically in Fig. 1. In analogy with Eqs. (5) and (11), the plastic velocity gradient tensor, Lp, can be defined as follows: Lp \u00bc ovp oxp \u00bc _KpK 1 p \u00fe KpX K 1 p \u00f025\u00de where vp and xp are the velocity and position vectors of the material particles in the \u2018\u2018P\u2019\u2019 configuration as shown in Fig. 1. Also, X* is the spin tensor associated with (QpR) rotation tensor, which represents the rigid body rotations from initial to \u2018\u2018P\u2019\u2019 configuration. Thus: X \u00bc \u00f0QpR\u00de \u00f0QpR\u00deT \u00bc _QpQT p \u00feQpXQT p \u00f026\u00de Inserting Eq. (26) into Eq. (25) and doing the following algebraic manipulations, we have: Lp \u00bc _KpQpQT p K 1 p \u00fe Kp _QpQT p K 1 p \u00fe KpQpXQT p K 1 p \u00bc \u00f0 _KpQp \u00fe Kp _Qp\u00deQT p K 1 p \u00fe KpQpXQT p K 1 p \u00bc \u00f0KpQp\u00de \u00f0KpQp\u00de 1 \u00fe \u00f0KpQp\u00deX\u00f0KpQp\u00de 1 \u00f027\u00de Using Eq. (21), Eq. (27) becomes: Lp \u00bc _CpC 1 p \u00fe CpXC 1 p \u00f028\u00de Eq. (28) represents an alternative form of Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000272_s0006-3495(76)85701-3-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000272_s0006-3495(76)85701-3-Figure4-1.png",
+        "caption": "FIGURE 4 Left: Beat pattern of the cilium of Opalina (courtesy of M. A. Sleigh). Right: Computed beat pattern of the model cilium.",
+        "texts": [
+            " The high magnification necessary for these small cilia, combined with their high frequency and the need for Nomarski illumination appears to preclude the taking of such cinemicrographs with present techniques, however. The photographs of Paramecium at raised viscosity published by Machemer (1972) did not have enough clarity and do not show a regular enough progression of the metachronal wave to make a construction of a ciliary cycle possible. Opalina. A tracing of the cilium of Opalina was taken from Sleigh (1960). This tracing, shown in Fig. 4 below, was constructed by Sleigh from photographs in an analogous way to that outlined above for Paramecium. Mytilus edulis was obtained from the Northeast Marine Specimens Co., Woods Hole, Mass. Gills of the mussels were excised and immersed in artificial seawater. The viscosity of the artificial seawater was raised by the addition of Ficoll to up to 330 cP. Moving films were made using a 45X Zeiss water immersion phase objective (Carl Zeiss, Inc., New York). For analysis large solitary cilia were selected on the films that were not obstructed in any way, showed a regular motion and stayed in focus over their whole length during the cycle",
+            " It appears that the stiffness generated by the model is approximately the correct one for Paramecium. ROBERT RIKMENSPOEL Contractile Events in Cilia 453 It should be noted that the recovery stroke shown in Fig. 3 (left) is much faster (duration 12 ms) than the one diagrammed by Sleigh (1968) and analyzed before (Rikmenspoel and Rudd, 1973), which had a duration of 25 ms. The value for the activation a,r,, found in the earlier analysis, which in the formulation of this paper would be a,,,, = 0.65, is therefore not in conflict with the value a,r,c = 1.25 presently reported. Opalina Fig. 4 (left) shows the tracing of the cilium of Opalina. The end of the effective stroke (beyond 225 ms in Fig. 4) is often slow and irregular, and it will not be considered here. The recovery stroke of Opalina was reproduced as shown in Fig. 4 (right) with an activation a,,, = 1.2. The stiffness, as generated by the model without corrections was found to give the correct maximum curvature in the bend in the cilium. During the effective stroke which starts at t = 150 ms, the cilium is S-shaped. In the model calculations, the motion of the tip of the cilium of approximately 0.15 ,um/ms could be reproduced with an activation a,,ff = 0.2. In this case the shape of the cilium did not show an inflexion, but it remained singly curved. The inflexion and the rather sharp curvature near the base at 150 and 175 ms could be obtained by having a second \"recovery\" stroke, but with the active moment opposite in sign from the original one, starting at t' = 125 ms and progressing with a velocity V' = 90 um/s. The activation of this second recovery stroke was a' = -0.2. Fig. 4 (right) shows the effective stroke thus computed, using the normal stiffness for the model. The general shape of the effective stroke is fairly represented in the model, but the curvatures in the model cilium are too small, especially near the tip. At 225 ms the curvature in the model is actually reversed from that in the data. Reducing the stiffness improves the shape of the computed position at 150 ms, but produces inflexions at the positions at 200 and 225 ms. It appears therefore that the general shape of the motion of the Opalina cilium can be explained with the model",
+            " The cilia of Opalina have been interpreted as an intermediate form of ciliary and flagellar beating patterns (Sleigh, 1968). Normal cilia show one effective and one recovery stroke. If Opalina with one effective and two recovery strokes is an intermediate between cilia and flagella, the logical extension would be that flagella have two effective and two recovery strokes. This proposal has indeed been made before (Rikmenspoel, 1971; Rikmenspoel and Rudd, 1973). Table IV summarizes the values for the essential quantities adopted in producing the motion of the model Opalina shown in Fig. 4 (right). Mytilus Fig. 5 shows the motion of the Mytilus cilium at viscosities of 1.4, 3.4, 12, 26, 103, and 330 cP. Two different cilia at 1.4 cP are displayed to give an impression of the range of variation between cilia. The motion of the Mytilus cilia is not completely regular, and the variation between different cycles of the same cilium is of the same magnitude as that between the two cilia at 1.4 cP. In the latter part of the effective stroke the angu- ROBERT RIKMENSPOEL Contractile Events in Cilia 455 lar velocity of the cilium is variable, with the cilium often coming to rest in an intermediate position (an \"interkinetic pause,\" Sleigh, 1968)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001087_1.3002329-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001087_1.3002329-Figure1-1.png",
+        "caption": "Fig. 1 A rigid sphere contacting an elastic half-space",
+        "texts": [
+            " The von Mises stress is defined as VM = 1 2 rr \u2212 2 + \u2212 zz 2 + rr \u2212 zz 2 + 6 r 2 + z 2 + rz 2 32 to the onset of yield at z /a=0.409 o that yielding occurs when VM= Y. It is noted that in axisymetric problems, the z-axis is a principal stress axis so that rr, , and zz along the z-axis are principal stresses. Furthermore rr= so that on the z-axis Eq. 32 becomes VM = rr \u2212 zz 33 nd in this case the von Mises yield condition is equivalent to the resca maximum shear stress criterion. In the three models conidered, a sphere indents an elastic half-space, as shown in Fig. 1. he applied load P represents PJKR, PDMT, or PGJ depending on he model chosen. 3.1 Subsurface von Mises Stress Fields. The subsurface von ises stress field contours for the JKR model on the verge of ielding are plotted using Eqs. 17 and 32 for =0.600 and JKR=4.90 in Fig. 2. It is noted that the stress field is singular at he boundary of the contact region, as is also true at the tip of a rack. Thus we only considered yield on the z-axis when using the KR theory. Similarly the subsurface von Mises stress field for the MT theory is plotted using Eqs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003812_robio.2012.6491108-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003812_robio.2012.6491108-Figure3-1.png",
+        "caption": "Fig. 3. Role of biarticular muscles, demonstrating relationship between GA, SO and VL.",
+        "texts": [
+            " All monoarticular muscles are relaxed when the apex is reached (close to 250 ms) to prepare for landing, where only the biarticular muscles absorb the impact. After impact, IL is once again regulated to reach the controlled value. A schematic demonstrating this activation pattern is shown in Fig. 2. To understand the role of biarticular muscles, we elaborate an explanation from [8]. Similarly to van Ingen Schenau, we can depict the body acting against an external force, such as the ground reaction force (GRF) during landing in Fig. 3. Notice that the total work done by the tip of the foot is the sum of works done by the other joints: \u0394W = Thip\u0394\u03b8hip +Tknee\u0394\u03b8knee +Tankle\u0394\u03b8ankle (2) Considering the frame of reference at the hip (Fig. 3B), the cartesian position of the foot would be defined as: x y = \u2212lthighcos(\u03b8hip)\u2212lshankcos(\u03b8hip\u2212\u03b8knee)\u2212l f oot cos(\u03b8hip\u2212\u03b8knee+\u03b8ankle) \u2212lthighsin(\u03b8hip)\u2212lshanksin(\u03b8hip\u2212\u03b8knee)\u2212l f oot sin(\u03b8hip\u2212\u03b8knee+\u03b8ankle) (3) and for small displacements the relation between the changes in joint angles with changes at the tip of the foot would be: [ \u0394x \u0394y ] = J \u23a1 \u23a3 \u0394\u03b8hip \u0394\u03b8knee \u0394\u03b8ankle \u23a4 \u23a6 (4) where J is the Jacobian, given by J = [ lthighsin(\u03b8h)+lshanksin(\u03b8h\u2212\u03b8k)+l f oot cos(\u03b8h\u2212\u03b8k+\u03b8a) \u2212lthighcos(\u03b8h)\u2212lshankcos(\u03b8h\u2212\u03b8k)\u2212l f oot cos(\u03b8h\u2212\u03b8k+\u03b8a) \u2212lshanksin(\u03b8h\u2212\u03b8k)\u2212l f oot sin(\u03b8h\u2212\u03b8k+\u03b8a) l f oot sin(\u03b8h\u2212\u03b8k+\u03b8a) lshankcos(\u03b8h\u2212\u03b8k)+l f oot cos(\u03b8h\u2212\u03b8k+\u03b8a) \u2212l f oot cos(\u03b8h\u2212\u03b8k+\u03b8a) ] (5) Finally, the relationship between a change in the external force and a change in joint torques would be: \u0394T = JT \u0394F (6) Considering Tankle as a reaction to GRF generated by ankle extensors, activating biarticular muscles would be more effective than relying on monoarticulars. The reason for this comes from the tapered muscular structure that mammals usually possess, where power (muscle) is concentrated on proximal members, connecting to distal members with fewer muscles or tendon/sheath structures, such as the human hand. Transferring energy from the strong knee extensor VL through GA reduces the force needed on SO (as demonstrated in Fig 3B). The relationship between SO and GA was quantitatively approached in [5]. We performed experiments by attaching the hindlimb on a linear slider, allowing the body to rotate along the sagittal plane. Due to horizontal translational constraints we added two slopes, each one 200 mm away from the slider\u2019s vertical projection, elevating 30 degrees to allow foot contact at steeper angles, as seen in Fig 3B. At the first phase of the experiment the pressure at IL, GA and ST was fixed at a determined value and the hindlimb hopped 25 times. The landing position of each jump was recorded and new parameters for IL, GA and ST were set. Instead of recording the landing position with angular information, we judged that adopting distance would be more ) precise, since the hindlimb\u2019s height is specially affected by the pressure at GA. For the second phase of our experiment we used data acquired from the first phase and controlled a hopping hindlimb, following a target value during 72 sequential jumps"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure9.5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure9.5-1.png",
+        "caption": "Fig. 9.5-1 Lateral vibration of a beam (see [9.2]). (This graph is reproduced with permission of the F. S. Tse family and co-authors.)",
+        "texts": [
+            " As a result, the solution has the form ( , ) sin ( ) cos ( ) x t D n x n ct n n = \u2212 \u2212 = \u2202 \u2211 2 1 2 2 1 21 L L (g) Th e fi rst condition of Eq. f yields D n x Tx J Gn pn sin ( )2 1 21 \u2212 = = \u2202 \u2211 L (h) Multiplying in Eq. h by sin(( ) / )2 1 2n x\u2212 L and integrating from x=0 to x=L, we obtain D T n J Gn p n= \u2212 \u2212 \u22128 2 1 1 2 2 1L ( ) ( )( ) (n=1,2,\u2026) (i) Th e torsional vibration of the rod is then given by ( , ) ( ) ( ) sin ( )( ) x t T J G n n x p n = \u2212 \u2212 \u2212\u22128 1 2 1 2 1 22 1 2 L L cos ( )2 1 2L1 n ct n \u2212 = \u2202 \u2211 (j) VibrationAnalysis_txt.indb 322 11/24/10 11:49:46 AM Continuous Systems | Chapter 9 323 9.5 Lateral Vibration of Beams Fig. 9.5-1 (a) shows a long, thin fl exible beam which can undergo displacement in the x-y plane. For simplicity, we begin our analysis with a beam whose displacement u is small compared with the beam length and is due to bending along the beam. Th e corresponding model is known as the EulerBernoulli beam [9.2]. Th e displacement u of the beam will depend upon both the position x along the beam and time t. Th us, u=u(x,t). Fig. 9.5-1 (b) shows a freebody diagram of a diff erential element of the beam. Th e fi gure itself shows the positive directions for the bending moment M and the shear force V. From a balance of forces and Newton\u2019s law, we have dx u t V dV V \u2202 \u2202 = \u2212 +( )+ 2 2 or simply \u2202 \u2202 = \u2212 \u2202 \u2202 2 2 u t V x , (9.5-1) where is the beam mass per unit length. Next, from a balance of moments in Fig. 9.5-1 (b), we have \u2202 \u2202 \u2212 =M x dx Vdx 0 or simply \u2202 \u2202 =M x V , (9.5-2) where M is the bending moment with positive direction defi ned as in Fig. 9.5-1 (b). VibrationAnalysis_txt.indb 323 11/24/10 11:49:48 AM Principles of Vibration Analysis 324 From elementary strength of materials analyses [9.5], we have the relation EJ u x M \u2202 \u2202 = 2 2 , (9.5-3) where EJ is the fl exural rigidity of the beam. Combining Eqs. 9.5-1 to 9.5-3, the diff erential equation for lateral vibration of the beam becomes \u2202 \u2202 = \u2212 \u2202 \u2202 \u2202 \u2202 \u239b \u239d\u239c \u239e \u23a0\u239f 2 2 2 2 2 2 u t x EJ u x (9.5-4) In the special case where the fl exural rigidity EJ is constant along the length of the beam, the above equation may be written as \u2202 \u2202 + \u2202 \u2202 = 2 2 2 4 4 0 u t a u x , (9",
+            " Th erefore, the modes of natural vibration and their frequencies will be established. Table 9.5-1 lists the common boundary conditions (see [9.2]). By superimposing all possible normal vibrations, the general expression for the free lateral vibration of beam becomes u x t A t B tn n n n n n ( , ) ( sin cos ),= + = \u221e \u2211 1 (9.5-17) where the constants An and Bn are to be determined by initial conditions. Example 9.5-1 Derive the frequency equation and eigenfunction for the lateral vibration of a uniform beam with the left end fi xed and the right end simply supported, as in Fig. 9.5-1 (a). Solution: Let the lateral displacement be u x t x q t( , ) ( ) ( )= . From Eq. 9.5-13, ( )x has the form ( ) (cos cosh ) (cos cosh ) (s x C x x C x x C = + + \u2212 + 1 2 3 in sinh ) (sin sinh ) x x C x x+ + \u22124 (a) Th e boundary conditions for the fi xed end are (1) u t( , )0 0= and (2) \u2202 \u2202 = = u x x 0 0 VibrationAnalysis_txt.indb 325 11/24/10 11:49:50 AM Principles of Vibration Analysis 326 Th ose for the simply supported end are (3) u L t( , )= 0 and (4) \u2202 \u2202 = = 2 2 0 u x x L Condition (1) requires that C1=0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001491_robot.2008.4543713-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001491_robot.2008.4543713-Figure1-1.png",
+        "caption": "Fig. 1. (a) Angle \u03b1 defined by the Law of Cosines of a triple. (b) Another derivation of angle \u03b1 by the constructing the triangle incircle of a triple.",
+        "texts": [
+            " We first give an important term, angle sum, in circle packing and its monotonic properties. We then define the circle packing problem. We prove that any complex K with fixed boundary radii always has a unique circle packing such that any two circles are tangent whenever their associated nodes form an edge in K. As the computation of a circle packing can be accomplished distributedly, it fits the needs of scalability for sensor networks. A. Angle Sum For each triple of radii ri, rj and rk, the Law of Cosines gives the angle \u03b1 in a corresponding triple of circles, as shown in Fig. 1(a). If we add these individual angles over the k triples involved, we get the angle sum \u03b8 for this label at v. Suppose },,,;{ k21v vvvvF L= is the flower for v in K. Vertex v belongs to m faces, where m=k if v is interior and m=k-1 if v is boundary. In a flower having central label r and petal labels { }k21 rrr ,,, L , the angle sum is given by the following summation formula, where m=k and rk+1=r1 if the flower is closed, and m=k-1 otherwise: \u2211 = + ++\u2212 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 ++ +\u2212+++ = m 1j 1jj 2 1jj 2 1j 2 j1 k21 rrrr2 rrrrrr rrrr ))(( )()()( cos),,,;( L\u03b8 (1) Note that the calculation of angle sum does not require any sensor node coordinates, only sensing radii are involved",
+            " LEMMA 1 (Monotonicity in flower). Let { }k21v vvvvF L,,;= denote a flower with central vertex v and a chain of k petals { }k21 vvv L,, . Let { }k21 rrrr ,,,; L be the list of radii of the circles in Fv and write )(v\u03b8 the angle sum at v as a function of k21 rrrr ,,,, L . Assume \u221e<r , )(v\u03b8 is strictly decreasing in central circle radius r and strictly increasing in petal radii ri that are finite. If 0r = , \u03c0\u03b8 kv \u2261)( . If \u221e=r , then 0v \u2261)(\u03b8 . It is not difficult to prove Lemma 1 by expressing the inradius rin (Fig. 1(b)) with the aid of the Heron's formula 1jj 1jj in rrr rrr r + + ++ \u22c5\u22c5 = (2) to obtain a easier derivative of the angle sum with a new expression. The flower of an interior node can be realized as a closed geometric flower if and only if \u03c0\u03b8 2= . We call \u03c0\u03b8 2rrrr k21 =),,,;( L the packing condition. LEMMA 2. Let ak21a rrrr \u03b8\u03b8 =:),,,;( L and bk21b rrrr \u03b8\u03b8 =:)',,',';( L be the angle sums of two flowers Fa and Fb of same number of petals. Suppose ba \u03b8\u03b8 = . If 'ii rr \u2265 for k1i ,,L= and 'ii rr > for some i (i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003080_jmer.9000033-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003080_jmer.9000033-Figure7-1.png",
+        "caption": "Figure 7. Dynamic model of engine in Adams/Engine methods.",
+        "texts": [
+            " Ts consist of fan, alternator, timing chain and oil pump resistance torque, as follow: AlternatorOilpumpgTifanS TTTTT +++= min (27) Ts approximately is 10-15% of total engine output torque. ffSCf JTTT \u03b1.\u2212\u2212= (28) For calculation of implied forces over relevant parts like crankshaft and connecting rod, a FORTRAN program was written used considering above equations. For checking the accuracy of program, the extract results of program for sample engine have been compared with results of Ranjbarkohan et al. 89 simulating the same engine in Adams/Engine software (Figure 7). The results of last section and Adams/Engine's output were compared. The compared results of two methods were shown in next figure. As mentioned, the main reason of failing of Nissan engine's crankshaft is using of downshifting in driving. It means shifting the gear from light gear (like 3) to heavy gear (like1) and usually is used for speed control of vehicle by the drivers in very steep roads with heavy loads. Before shifting from 3 to 1: 3333 3 3 3 3 3 . ggce g c g e g rr \u03c9\u03c9\u03c9 \u03c9 \u03c9 \u03c9 \u03c9 ==\u21d2== (29) After disengagement of engine and shifting from 3 to 1 and before releasing the clutch: (30) where, rg1 is the first gear ratio, rg3 is the third gear ratio, rd is the differential ratio, \u03c9c1 is the clutch plate rotational velocity in gear 1, \u03c9c3 is the clutch plate rotational velocity in gear 3, \u03c9g1 is the transmission rotational velocity in Gear 1, \u03c9g3 is the transmission rotational velocity in Gear 3, \u03c9e1 is the engine rotational velocity in Gear 1 and \u03c9e3 is the engine rotational velocity in Gear 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001763_s11666-010-9534-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001763_s11666-010-9534-8-Figure1-1.png",
+        "caption": "Fig. 1 Laser cladding under high-frequency microforging. (a) Mechanism of microforging. (b) Layout of high-frequency microforging",
+        "texts": [
+            " The results showed that high-power ultrasonic microplastic deformation forging could dispel the surface tensile stress and form compressive residual stress. The friction coefficient was noticeably reduced. Surface hardness, wear-resisting performance, roughness, and fatigue strength, and so forth all improved. Nanometer layer of about 100 lm depth came into being on the metal surface, which has been applied in practice. Based on mechanical effects on metal surface produced by high-frequency microforging process, this paper tries a new method to regulate the cracking behavior and friction and wear characteristics (as Fig. 1 shows). The temperature of the forging zone on the laser cladding layer was kept between 750 and 900 C using mechanical equipment. Under highfrequency microforging, tiny plastic deformation develops on the layer, and its organization turned from as-cast to as-forge, and the weak interface began to be welded; meanwhile, compressive plastic deformation eliminated tensile strain on the layer and formed compressive residual stress, which offset the tensile stress because of the different coefficients of thermal expansion (CTE) between the cladding and the substrate during subsequent cooling and thus reduced the surface tensile stress to the minimum"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure4.19-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure4.19-1.png",
+        "caption": "Figure 4.19 Vectors to point \u2018q\u2019",
+        "texts": [
+            " Since Laplace transforms provide a complete solution to linear differential equations and their stimulus we can obtain both the transient response to any function of time (provided that we can define the input time function as a Laplace transform) and also the frequency response by substituting s = j into the system transfer function. Suppose we have a system defined by the Laplace transform: F s = k s\u2212z1 s\u2212z2 s\u2212zm s\u2212p1 s\u2212p2 s\u2212pn At any point q in the \u2018s\u2019 plane we can determine the value of F s by the previously described expression which means: F s = k product of vectors from the zeros to point q product of vectors from the poles to point q This point is illustrated graphically by Figure 4.19 which shows an arbitrary point q in the \u2018s\u2019 plane with the vectors from each zero and pole in a typical system. The vector lengths are multiplied together and the vector angles summed to obtain the vector value of F s at the selected point q. Suppose now that we select point q to be on the imaginary axis of the \u2018s\u2019 plane. Here s= j and so F s will describe the frequency response of the system at that specific frequency. Thus moving along the j axis and calculating the ratio of the vector products in the above expression for F s generates the frequency response of the system defined in the \u2018s\u2019 plane"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002806_iceee.2011.6106621-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002806_iceee.2011.6106621-Figure2-1.png",
+        "caption": "Fig. 2: Inertial and Body frame of the four-rotor aircraft.",
+        "texts": [
+            " This type of counter-rotating rotor-pairs setup will cancel out the rotating force produced by the rotating rotor where the conventional helicopter is using the tail rotor to counter it. The flight direction control of the Quad-Rotor Flying Robot is described below. The front and rear motors rotate counterclockwise (CCW), while the other two motors rotate clockwise (CW). Gyroscopic effects and aerodynamic torques tend to cancel in trimmed flight. The main thrust is the sum of the thrusts of each motor, the pitch torque is a function of the difference f1 \u2212 f3, the roll torque is produced by the difference f2 \u2212 f4, and the yaw torque is the sum of \u03c4M1 + \u03c4M2 + \u03c4M3 + \u03c4M4 (see Figure 2), where \u03c4Mi is the reaction torque of motor i due to shaft acceleration and the blade\u00b4s drag. Let \u0393I = {iI , jI , kI} be the inertial frame, \u0393B = {iB , jB , kB} denote a set of coordinates fixed to the rigid aircraft as is shown in Fig. 2. Let q = (x, y, z, \u03d5, \u03b8, \u03c8)T \u2208 R6 = (\u03be, \u03b7)T be the generalized coordinates vector which describe the position and orientation of the flying machine, so the model could be separated in two coordinate subsystems: translational and rotational. They are defined respectively by \u2022 \u03be = (x, y, z)T \u2208 R3: denotes the position of the aerial vehicle\u00b4s center of mass relative to the inertial frame \u0393I . \u2022 \u03b7 = (\u03d5, \u03b8, \u03c8)T \u2208 R3: describe the orientation of the aerial vehicle and (\u03d5,\u03b8,\u03c8) are the three Euler angles: roll, pitch and yaw, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002546_s1052618810050018-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002546_s1052618810050018-Figure2-1.png",
+        "caption": "Fig. 2. Mechanism with four degrees of freedom and two parallelograms: (a) scheme of the mechanism; (b) system of kinematic screws in the usual configuration; (c) system of kinematic screws in the special provision.",
+        "texts": [
+            " Here, the linear motors provide the position of the output link with constant ori e14y 0 e22x 0 e22z 0 e23x 0 e23z 0 e24x 0 e24y 0 e33x 0 e33y 0 e34x 0 e34y 0 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 39 No. 5 2010 DEVELOPMENT OF MECHANISMS OF PARALLEL STRUCTURE 409 entation, and the rotational engine provides orientation of output link\u2014rotation around the vertical axis. The gear ratio with linear movement and rotation is equal to unity. Consider the mechanism in which three kinematic chains are imposed on two connections (Fig. 2). The first and second kinematic chains are composed of one drive rotational pair (rotary engine) located on the base, one intermediate rotational pair located with the axis parallel to the axis of the rotary engine, and a cylindrical two movable pair (the axes of the finite cylindrical pairs of the two chains are the same). The third kinematic chain, as in the previous case, contains one rotational drive pair installed on the base, one trans lational drive pair (the axes of these two pairs are the same), and two translational pairs made in the form of hinged parallelograms (Fig. 2a). Single screws characterizing the position of the axes of these kinematic pairs have the coordinates E11(0, 0, 1, , , 0), E12(0, 0, 1, , , 0), E13(0, 0, 1, , , 0), E41(0, 0, 0, 0, 0, 1), E21(0, 0, 1, , , 0), E22(0, 0, 1, , , 0), E23(0, 0, 1, , , 0) = E13(0, 0, 1, , , 0), E24(0, 0, 0, 0, 0, 1) = E14(0, 0, 0, 0, 0, 1), E31(0, 0, 1, 0, 0, 0), E32(0, 0, 0, 0, 0, 10), E33(0, 0, 0, , , 0), E34(0, 0, 0, , , 0). Screws E14, E24, E32, E33, and E34 have an infinitely large parameter. The rest of the screws have a zero parameter. All the kinematic chains imposed on the two links can be considered as repetitive; they deter mine the number of degrees of freedoms equal to four. Power screws of connections imposed by the kine matic chains, as in the previous case, have the coordinates (Fig. 2b) R1(0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0). Accordingly, all the kinematic screws of movement of the output link can be represented as screws mutual to the power screws \u21261(0, 0, 0, 1, 0, 0), \u21262(0, 0, 0, 0, 1, 0), \u21263(0, 0, 0, 0, 0, 1), \u21264(0, 0, 1, 0, 0, 0); screws \u21261, \u21262, and \u21263 have an infinitely large parameter; screw \u21264 has a zero parameter. Special provisions relating to the loss of one or more degrees of freedom arise if the kinematic screws corresponding to the unit vectors Ei1, Ei2, Ei3 (i = 1, 2) or to the unit vectors E33, E34 are linearly dependent. This occurs if any three screws Ei1, Ei2, Ei3 (i = 1, 2) fall in one plane or if the two screws E33, E34 are par allel. In particular, if any three screws Ei1, Ei2, Ei3 (i = 1, 2) fall in one vertical plane, there exist three power screws imposed by kinematic chains R1(0, 0, 0, 1, 0, 0), R2(0, 0, 0, 0, 1, 0), and R3(1, 0, 0, 0, 0, 0) (Fig. 2c) and only three kinematic screws of the movement of the output link mutual to these screws \u21261 (0, 0, 0, 0, 1, 0), \u21262(0, 0, 0, 0, 0, 1), and \u21263(0, 0, 1, 0, 0, 0). Note that R3 is located on the axis x. e11x 0 e11y 0 e12x 0 e12y 0 e13x 0 e13y 0 e21x 0 e21y 0 e22x 0 e22y 0 e23x 0 e23y 0 e13x 0 e13y 0 e33x 0 e33y 0 e34x 0 e34y 0 410 JOURNAL OF MACHINERY MANUFACTURE AND RELIABILITY Vol. 39 No. 5 2010 GLAZUNOV et al. If all drives are fixed, then, as in the previous case, there are six force screws imposed by kinematic chains R1, R2, R3, R4, R5, and R6"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002006_j.cirp.2010.03.071-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002006_j.cirp.2010.03.071-Figure1-1.png",
+        "caption": "Fig. 1. Prototyping deep drawing tool for stainless steel sheets (basic body St52-3, radii clad with Cu85Al10Ni5).",
+        "texts": [
+            " These alloys feature a high surface hardness as well as low friction coefficients. Relevant components of conventional tool versions, e.g. deep drawing of stainless steel sinks, are entirely made up of the material AMPCO 25. This leads however, to high costs for semi-finished parts and the need for special process steps. The implementation of direct laser deposition raises the possibility of using a low-cost material (e.g. 1.0570/St52-3) for the basic body of the tool and a limitation of the coating (coat thickness 2\u20134 mm) to relevant areas [1,2]. Fig. 1 illustrates the application of a prototyping tool for deep drawing of stainless steel sinks. Here the blankholder as well as the drawing die were entirely coated, while the coating of punch and die was restricted to the relevant areas. The challenge in manufacturing these coated tools arises from cladding three-dimensional tool surfaces. The quality of the clad must be constant at any surface orientation. Due to the shape of the tool surface the metal deposition has to be carried out in 2 dimensions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000966_la802967p-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000966_la802967p-Figure5-1.png",
+        "caption": "Figure 5. Capillary rise height h (dashed line) and the decay length l\u221e (solid line), as a function of the attack angle \u03c6 of the uniform director field. (a) \u03c9 \u2261 wIN/\u03b3IN ) 0.5. (b) \u03c9 ) -0.5. Complete wetting is presumed.",
+        "texts": [
+            " 1, the solution to eq 3 for the interface profile y(x) \u223c exp(- x/l\u221e) turns out not to depend on the angle \u03d1, with decay length l\u221e ) lc \u221a2 \u221a1+\u03c9 cos2 \u03c6 (6) This shows that the profiles for isotropic fluids with \u03c9 ) 0 and decay length l\u221e) lc/ 2 decay more rapidly than those of nematic fluids if \u03c9 > 0 and less rapidly if \u03c9 < 0. We are also able to conclude from eqs 5 and 6 that the rise height and the decay length are not symmetric for the cases \u03c9 and -\u03c9, so biases for homeotropic or planar anchoring produce quite different interface profiles. In fact, the rise height and the far-field decay depend in very different ways on the angle \u03c6, as shown in Figure 5. This means that the profiles for different attack angles \u03c6, presented in Figure 4, must intersect if plotted in the same figure, as in fact they do. (iii) Contact Angle. From the boundary term we obtain by functionally minimizing the free energy, eq 3, we extract an implicit equation for the contact angle \u03d1 ) -arctan(1/y\u2032(0)). It reads cos \u03d1+\u03c9[cos \u03d1 sin2(\u03d1- \u03c6)+ (sin \u03d1- sin2 \u03d1) sin(2\u03c6- 2\u03d1)]+ k) 0 (7) where k \u2261 (\u03b3NW + wNW sin2 \u03c6 - \u03b3IW)/\u03b3IN. Note that eq 7 for the contact angle is also asymmetric for the cases of planar (\u03c9 < 0) and homeotropic alignment (\u03c9 > 0)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001115_j.triboint.2009.11.005-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001115_j.triboint.2009.11.005-Figure2-1.png",
+        "caption": "Fig. 2. Test configuration: (a) flat on flat configu",
+        "texts": [
+            " This characterization was carried out by means of uniaxial tensile and compression tests on a universal test machine at quasi-static speed (10 mm/min). As will be explained in the following sections, the experimental data will be used for both input and validation of the hyperelastic model of the rubber in FE simulation. The tribometer test configurations selected are the flat on flat and the flat on cylinder geometries, both run with reciprocating motion of the rubber specimen. The material system described above is tested under these two configurations in dry conditions (no lubricant or grease): Flat on flat (Fig. 2a). Flat on cylinder (Fig. 2b). The above detailed tribometer configurations use an encapsulated fixing tooling for the rubber specimen. This steel tooling fixes a rubber cylinder of 12 mm diameter and 6 mm height to the reciprocating arm of the tribometer, using no glue or adhesive, and leaving 2 mm of \u2018\u2018free height\u2019\u2019 of rubber in the contact zone with the countermaterial to avoid contact between the tooling itself and the countermaterial when the rubber deforms under the vertical load. In addition, this configuration avoids excessive deflection of the rubber when sliding and allows a distribution of contact pressures more uniform if compared to other fixing toolings [21]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001069_oca.813-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001069_oca.813-Figure1-1.png",
+        "caption": "Figure 1. (a) Single-link flexible-joint robot system and (b) flexible-joint system.",
+        "texts": [
+            " Besides, the flexible joint robot is more close to the real robotic system. According to the simulation results, the input estimation algorithm is an effective observer for estimating the disturbance torque input, and the LQG controller can effectively cope with the situation that the disturbance exists. Finally, higher control performance of the combined method for joint control of the robotic system can be further verified. Consider the single-link flexible-joint robot system that is driven by the DC servomotor as shown in Figure 1(a). The mathematical equations are shown as follows: Jl\u0308l(t) + Bl\u0307l(t) + ks[ l(t) \u2212 m(t)] + MgL2 sin l(t) + Fd(t)L1 = 0 (1) Copyright q 2007 John Wiley & Sons, Ltd. Optim. Control Appl. Meth. 2008; 29:101\u2013125 DOI: 10.1002/oca Jm\u0308m(t) + Bm\u0307m(t) \u2212 ks[ l(t) \u2212 m(t)] = kt Iq (2) La I\u0307q(t) + Ra Iq(t) + ke\u0307m(t) = vt (t) (3) where Jm is the motor-side inertia, Jl is the load-side inertia, ks is the torsional spring constant, M is the link mass, L1 is the length of link, L2 is the distance to the centre of mass of link, Fd is the external force, kt is the motor torque constant, ke is the emf constant, Iq is the current, La is the inductance, Ra is the resistance, vt is the control input (voltage) of motor, m is the motor-side angular displacement, Bm is the motor-side viscous coefficient, Bl is the loadside viscous coefficient, and l is the link angular displacement. Furthermore, in Equation (2), kt Iq = m represents the motor torque. Equation (1) reveals that the single-link flexible-joint robot is a nonlinear system. When considering the load torque l(t) = MgL2 sin l(t) and the external force torque d(t) = Fd(t)L1 for the joint system as shown in Figure 1(b), Equation (1) can be rewritten as follows: Jl\u0308l(t) + Bl\u0307l(t) + ks[ l(t) \u2212 m(t)] = \u2212 ld( l, t) (4) where ld( l, t) = l(t)+ d(t) is the disturbance torque input. Let x1 = l, x2 = \u0307l, x3 = m, x4 = \u0307m and x5 = Iq . When considering the system process noise (system uncertainty or model error), the system equation can be rewritten as state-space description: X\u0307(t) =AX(t) + Bu(t) + Gw(t) (5) Z(t) =HX(t) + v(t) (6) Copyright q 2007 John Wiley & Sons, Ltd. Optim. Control Appl. Meth. 2008; 29:101\u2013125 DOI: 10",
+            " For = 1, the above algorithm reduces to that of the conventional sequential least-squares algorithm, which is appropriate only for a constant input estimation. The smaller the , the faster the forgetting efficacy will be. Nevertheless, the rapid forgetting efficacy reduces the smoothing efficacy. For this reason, it has to consider the interaction of smoothing and forgetting while deciding the forgetting factor . The block diagram of recursive Copyright q 2007 John Wiley & Sons, Ltd. Optim. Control Appl. Meth. 2008; 29:101\u2013125 DOI: 10.1002/oca IE algorithm is shown in Figure 1, and the procedure for estimating the system inputs using the IE approach is summarized in the following [16]: (i) Derive and identify the system state-space model, i.e. Equations (7) and (8), and measure the system responses X(k). (ii) Use the KF equations, i.e. Equations (9a)\u2013(9g), to obtain the innovation covariance s(k), innovation Z(k), and Kalman gain Ka . Copyright q 2007 John Wiley & Sons, Ltd. Optim. Control Appl. Meth. 2008; 29:101\u2013125 DOI: 10.1002/oca (iii) Use the recursive least-squares algorithm, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000222_0045-7930(74)90007-3-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000222_0045-7930(74)90007-3-Figure1-1.png",
+        "caption": "Fig. 1. Two spheres moving arbitrarily.",
+        "texts": [
+            " A single sphere moving through an unbounded fluid otherwise at rest is described by the velocity potential, tp(r) : - ~ [U\" (r -- ;)], (1) where a is the sphere radius, U is the sphere velocity (a vector quantity), ~ is the vector distance to the sphere center from the origin of coordinates, and r ( - I r l) is the distance from the origin of coordinates to the point of interest. Suppose that a second sphere is somewhere in the infinite fluid field and moves in an arbitrary direction as depicted in Fig. 1. Each sphere will, to some degree, influence the fluid near the other. The extent of this interaction depends on the distance between the spheres, their velocities, and their radii. For two spheres far enough away from one another that there is negligible interaction, the velocity potential for each sphere is 1 ( ~ 13[Un-(r - - ~n)] (2) qg,(r) = - ~ \\1 r _ %,1/ where the index n indicates the sphere under consideration. This means that the potential for each sphere is the same as if only one sphere were present"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001559_acemp.2007.4510568-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001559_acemp.2007.4510568-Figure1-1.png",
+        "caption": "Fig. 1. Homopolar hybrid excitation synchronous machine",
+        "texts": [
+            " In the established models, ferromagnetic parts reluctance and the lamination effect has been taken into account. However, the saturation effect and leakage flux has been neglected. The validation of those models has been established by comparing them to the finite-elements simulations and the obtained results are satisfying. The studied machine is an usual flux concentration machine with two additional stator excitation coils as well as the necessary armatures for the canalisation of the flux they will create (Fig.1). These armatures are ferromagnetic massive parts while the stator and the rotor of the machine are laminated in order to limit the eddy currents. The structure detailed in this article has six pairs of poles and distributed coils. In the case of a homopolar configuration, the circuit of the double excitation acts only on one pole. The two double excitation winding currents have the same values and opposite directions [2]. The developed model is based on the main flux path and a part of the magnets leakage flux\u2019s path. In this paper, a 3D magnetic equivalent circuit is proposed and its validity is established by comparing the results it gave to those given by a finite-element model. In addition to the polar symmetry, the homopolar structure has a plan of symmetry. This plan is noted xOy and illustrated in Fig.1. Consequently, twelfth of the structure is enough in order to study the machine\u2019s behaviour (Fig.2). A. 2D flux path 2D flux path consists of the path of the part of the flux created by permanent magnets that participate to the energy conversion (Fig. 3). Its modelling is shown in Fig. 4. Reluctances appearing in the model are indicated in the nomenclature. Fig. 4. Magnetic equivalent circuit for 2D\u2019s path Where : a\u03b5 is the magnetomotive force produced by permanent magnets : 0\u00b5 \u03b5 aerB a \u00d7 = ea is the permanent magnet thickness Br is the residual flux density The total excitation flux is given by : ( ) ( )( ) '22' ' 2 1212341212 1212 RRRRRRR RR ++++ + \u00d7= eqa a v p \u03b5 \u03d5 (1) The reluctances appearing in the previous expression are calculated from the dimensions and the magnetic permeability of each part of the machine [7] and [8] : ( )\u222b \u00d7 = 2 1 l l lS dl \u00b5 R (2) Where : \u00b5 is the magnetic permeability of the concerned part of the circuit l is the length of the flux\u2019s path S is the cross-sectional area of the path The calculation of the reluctance of a half rotor pole is given as an exemple"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.25-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.25-1.png",
+        "caption": "Fig. 1.25. Large disturbance in a curve. New initial state vector (Av, Ar) after the action of a lateral impulse S. Once outside the domain of attraction the motion becomes unstable and may get out of control.",
+        "texts": [
+            " As a result, all trajectories starting above the lower separatrix tend to leave the area. This can only be stopped by either quickly reducing the steer angle or enlarging 6 to around 0.2rad or more. The latter situation appears to be stable again (focus) as has been stated before. For the understeered vehicle of Fig. 1.24 stability is practically always ensured. For a further appreciation of the phase diagram it is of interest to determine the new initial state (ro, Vo) after the action of a lateral impulse to the vehicle (cf. Fig. 1.25). For an impulse S acting at a distance x in front of the centre of gravity the increase in r and v becomes: S x S A r - , A v - (1.98) I m which results in the direction a A r x a b ,dv b k 2 (1.99) TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 51 The figure shows the change in state vector for different points of application and direction of the impulse S (k 2 = I/m- ab). Evidently, an impulse acting at the rear (in outward direction) constitutes the most dangerous disturbance. On the other hand, an impulse acting in front of the centre of gravity about half way from the front axle does not appear to be able to get the new starting point outside of the domain of attraction no matter the intensity of the impulse"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001725_pime_conf_1967_182_019_02-Figure14.16-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001725_pime_conf_1967_182_019_02-Figure14.16-1.png",
+        "caption": "Fig. 14.16. Rolling four-ball configuration",
+        "texts": [
+            " Using a radio-activated steel element, which rolls against inactive steel, and performing autoradiography on the inactive elements, it is found that strongly held transfer particles occur only if the lubrication is by a small quantity of oil such that there is an opportunity for loose wear particles to be re-deposited and subsequently rolled into the surface. If there is copious lubrication which will remove loose particles before they are rolled over, then no strongly held transfer particles are found. The wear phenomenon can, accordingly, be described by giving the wear rate as a function of the operating conditions. In references (12) and (40)~ experiments are cited in which the wear rate was determined for a number of mineral oils and synthetic fluids used to lubricate balls rolling on each other in the configuration illustrated in Fig. 14.16. The wear rate of only one (activated) ball was measured, using scintillation counting of the lubricant and of washings from the test configuration. A run-in period exists if the test is started with new balls, and thereafter the wear rate becomes steady. The experiments were conducted in the partial e.h.d. range, and in this range, the Vol I82 P t 3A at UNIV NEBRASKA LIBRARIES on September 8, 2015pcp.sagepub.comDownloaded from ROLLING CONTACT FAILURE C( equilibrium wear rate can be expressed as a function of the following variables : (a) the real area of asperity contact; (b) the slide-roll velocity ratio",
+            " For higher values of to z 3, the wear rate is about 100 times lower and, at these levels, it is negligible as a failure mode. The wear law cited states that the wear constant K is (as a first approximation) independent of the e.h.d. condition, and characterizes wear properties (boundary lubrication properties) of the lubricant-material combination of the rolling contact for the specific geometry and kinematic situation of a test. Table 14.5 shows some typical wear constants obtained in the four-ball configuration of Fig. 14.16 for a number of often used rolling bearing lubricants. The numbers given are only order of magnitude estimates, but they suffice to illustrate the difference between lubricants. The effect of the magnitude of sliding on wear rates seems to depend on the e.h.d. condition. With an almost complete e.h.d. film, the effect of sliding on wear seems to be minor. When there is less film and more substantial asperity contact, the wear rate increases by a factor of 10 and more as a ball-to-ball configuration is made to roll with a high amount of sliding (typical of high contact angle ball bearings) instead of zero sliding"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001341_j.mechmachtheory.2008.02.015-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001341_j.mechmachtheory.2008.02.015-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of the axes of a universal joint and a leg.",
+        "texts": [
+            "onsidering the novelty of the method and creativity of the research that is detailed in the two papers [Mechanism and Machine Theory, Vol. 33, No.7, pp. 993\u20131012, 1998 and Vol. 33, No.8, pp.1135\u20131152, 1998], we believe that correcting the following errors appearing in these two papers would enhance their valuable contribution to this field of research. 2008 Elsevier Ltd. All rights reserved. Errors and the proposed corrections Error 1: In [1,2] the moment of each universal joint is considered to be along the direction of its corresponding leg. For a universal joint with revolute joints in the k\u0302 and y\u0302 directions, as shown in Fig. 1, there is no restriction moments in those directions; the notation x\u0302 is used for unit vector in the x direction. Therefore the universal joint can only impose a moment in the h\u0302 direction, where h\u0302 is the normal vector of the plane formed by the vectors k\u0302 and y\u0302, that is h\u0302 \u00bc k\u0302 y\u0302 \u00f01\u00de Thus instead of writing the moment of the universal joint Muniversal as Muniversal \u00bc Mz\u0302 \u00f02\u00de where M is the value of the moment of the universal joint and z\u0302 is the unit vector along the direction of the leg, Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002824_2012-01-0908-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002824_2012-01-0908-Figure5-1.png",
+        "caption": "Figure 5. Transition of oil film pressure inside pin boss",
+        "texts": [
+            " The graph in the lower part of Figure 4 shows the correlation between the up-and-down movements on an actual engine and in the simulation, which were compared in order to verify the difference in pin behaviors depending on oil film width. What is notable is that, with W2, which is equal to the pin hole width inside of the oil groove, the highest correlation was observed with the behavior on an actual engine. This suggests a high possibility that, when the pin hole has an oil groove, the area outside the groove is small and has low oil film pressure, which leads to a smaller effect of shock absorbing the impact between the pin and the pin hole and then to worsened pin noise. Figure 5 shows the results of an analysis of the transition of oil film pressure inside the pin hole. The color contours in the upper part indicate the oil film pressure at, from left to right, 25 deg. BTDC, 23 deg. BTDC, and 15 deg. BTDC. It was confirmed that, as the piston pin moves in a circle along the surface of the pin hole, the oil film pressure increases in a circle and, when the pin passes the side relief on the thrust side and reaches the upper part of the hole (15 deg. BTDC), the pressure rapidly goes up. As shown in the diagrams in the lower part of Figure 5, this indicates that, when the pin circles and reaches the upper part of the pin hole, the oil film pressure increases rapidly, possibly causing oil to leak from the side relief area on the anti-thrust side. It proved highly possible that this lowered the oil film pressure at the upper part of the pin hole, decreasing the effect of shock absorbing the pin impact against the pin hole and worsening pin noise. Oil leak from the side relief on the anti-thrust side was also observed in the visualization of lubricant leak on an actual engine[1], and this was confirmed by this simulation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001219_s12541-009-0102-4-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001219_s12541-009-0102-4-Figure8-1.png",
+        "caption": "Fig. 8 Load application points and displacement measurement points for estimating the normal, lateral, and axial compliances of the Z-axis feed system",
+        "texts": [],
+        "surrounding_texts": [
+            "A virtual prototype of the ultra-precision machine for machining large-surface micro-features, which was constructed based on ANSYS software to estimate its compliances, is presented in Fig. 10. The virtual prototype was composed of 134,729 nodes, 534,911 solid elements, and 408 matrix elements. The matrix elements were introduced in order to represent the normal and lateral stiffnesses of the hydrostatic guideways, the radial and thrust stiffnesses of the hydrostatic bearings, and the axial stiffness of the linear motors. As the boundary condition for the structural analysis and measurement, the movement of the ultra-precision machine was restricted in the vertical direction at the four supporting points of the bed, as shown by the red arrows in Fig. 10. The bed, column, cross beam, and feed tables were made of cast iron (GC300), the C-axis shaft and bracket were made of steel (SCM440, SS400), and the mover/rotor and stator of motors were made of Fe-Si. Table 3 shows the material properties used for the virtual prototype."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000001_1.2032993-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000001_1.2032993-Figure5-1.png",
+        "caption": "Fig. 5 FE mod",
+        "texts": [
+            " To take the full advantage of friction force in power transmission, it is desirable to have the actual operating friction coefficient ofc set close to, but just slightly smaller than, the maximum available friction coefficient max at the frictional contacts. Since the operating friction coefficient ofc of the friction drive can be adjusted by altering the support to contact stiffness ratio for a given geometrical configuration, this offers a practical and very useful means for design optimization. To quantify the stiffness ratio, providing a practical guide to support stiffness selection, a 2D plane stress finite element model was constructed in ABAQUS 1 to study the wedge loading action. As seen in Figure 5, the outer ring and the output shaft was simplified as a thick section ring connected to a thin web, which, in turn, was connected to a center node using ABAQUS MPC beam elements. The sun roller was simplified as a 2D analytical rigid body. The loading planet was modeled as a ring with contact elements set 1 ABAQUS is a registered trademark of ABAQUS, Inc. It is a finite element analysis software. OCTOBER 2005, Vol. 127 / 859 data/journals/jotre9/28735/ on 02/26/2017 Terms of Use: http://www.asme"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003031_iros.2011.6094957-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003031_iros.2011.6094957-Figure1-1.png",
+        "caption": "Fig. 1. Workspaces for the Barrett Hand: a) Reachable points for the fingertips; b) Set of points \u03a6fc.",
+        "texts": [
+            " To compute \u03a6fc, the hand configuration space is uniformly sampled, and the position of the fingertip region is computed via the direct kinematics of the hand. The FC condition can be tested using the normals to the fingertips, a predefined coefficient of friction \u00b5p, and a common coordinate system with its origin located at the centroid of the considered fingertip points. The artificially imposed \u00b5p should be the maximal value expected for the applications of the real handobject system. As an example, Fig. 1a shows for a Barrett hand [21] the set of reachable points for a patch defined on each fingertip. Fig 1b shows for the same hand the set \u03a6fc, computed with a friction coefficient of \u00b5p = 0.5. Note that in this case the set \u03a6fc is composed by three different subsets of points, hereafter called \u03c6i, one for each fingertip (\u03a6fc = \u03c61 \u222a \u03c62 \u222a \u03c63). The computation of the graspability map requires the following data: \u2022 A 3D object model (with the assumptions described in Section III-A). \u2022 A set \u03a6fc of reachable points for the fingertips of a mechanical hand, which potentially lead to force closure precision grasps (Section III-B)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000414_bfb0119398-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000414_bfb0119398-Figure4-1.png",
+        "caption": "Figure 4. The new Sarcos Treadport.",
+        "texts": [
+            " It is possible that the tether force changes the biomechanics of walking also, because a subject has to lean forward more. Thus a comparison of the knee-hip cyclograms for tether force versus slope walking would be of interest. The results on slope perception of 0.5 degrees while walking also point to a high sensitivity to slope while walking. While tether force can substitute for treadmill tilt, it would be desirable also to have a fast tilt mechanism. Consequently a second-generation Treadport has been designed, with one of its key features a fast tilt mechanism (Figure 4). The specifications on tilt are ~:20 degrees in 1 second. In the first-generation Treadport, the rotation occurs at the back of the platform for upward tilt from horizontal, whereas for downward tilt from horizontal the rotation occurs in the middle of the platform. The unfortunate result for the upward tilt is that the user is lifted as well as tilted. In the new Treadport, the tilt axis is fixed in the middle of the platform. The platform has also been made larger (6x10 feet) to allow more freedom of motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000837_6.2008-7192-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000837_6.2008-7192-Figure1-1.png",
+        "caption": "Figure 1. A SensorCraft configuration.",
+        "texts": [
+            " An ISR platform like SensorCraft has very limited maneuvering requirements; it needs only enough maneuvering envelope to take off and land. The turbulence loadings encountered exceed the expected maneuver loads so the gust loadings become the critical design case. Gust load alleviation applied to the SensorCraft configuration is an effort to use the structure in an efficient manner during gust loading and for an aerodynamic shape that is very efficient at 1 g cruise. The configuration of a possible SensorCraft is shown in Fig. 1. The requirements of aerodynamic efficiency along with low observable signature are balanced with a flying wing with a trapezoidal body with long outboard wings. The aerodynamic and structural loads tend to be concentrated at the interface between the outboard wing and the main body which will be denoted as the wing break. The entire load generated by the outboard wing has to be 2 Engineer 3, Vehicle Systems, 1 Hornet Way, M/S 9V21/W6. 3 Engineer 3, Vehicle Systems, 1 Hornet Way, M/S 9V21/W6. T 26th AIAA Applied Aerodynamics Conference 18 - 21 August 2008, Honolulu, Hawaii AIAA 2008-7192 Copyright \u00a9 2008 by Northrop Grumman Corporation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001786_0731684409357800-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001786_0731684409357800-Figure3-1.png",
+        "caption": "Figure 3. Schematic of flexible composite skin structure.",
+        "texts": [
+            " After checking the airproof performance of the system, the chamber is vacuumed again. When the baffle valve is inside, allow the helium gas to enter into the high pressure chamber with 1 atm, and measure the pressure at the measured chamber for almost 10 h. Finally the p t curve is processed to obtain the leakage parameters. The intact flexible composite refers to a film structure textile which uses fiber textile as its reinforcedmaterial and thermal-plastic polymer as its matrixmaterial. Currently, flexible composite shown in Figure 3 is named as skin structure in aerostats. This skin structure consists of three different parts, i.e., protected layer, load carrier, and helium barrier, which have different structures and compositions as shown in Table 1. The main compositions of helium barrier (Tedlar film) and load carrier are TEDLARandVectran/PU adhesive, respectively. Themajor materials of protected layer are TPU intermingling with inorganic nano-particles TiO2. Figure 4 shows the typical measurement specimens: the intact specimen, the inner damage specimen, and the surface damage specimen"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001034_cae.20326-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001034_cae.20326-Figure9-1.png",
+        "caption": "Figure 9 Machining process simulation driven by a NC program: (a) turning process simulation and (b) milling process simulation. [Color figure can be viewed in the online issue, which is available at wileyonlinelibrary.com.]",
+        "texts": [
+            " Moreover, the NC program should be edited according to the tool setting and tool compensation data. Figure 6b shows the view of a NC program ready to run. With machining process simulation, the material removal process and the movement of a machine tool are simulated visually with real-time animation. Therefore, the relationship between the NC instructions (and operations) and their executed results can be soundly established in a short period. It is helpful to speed up the training progress. Figure 9 shows the simulation of turning process and milling process driven by NC programs, respectively. In CNC Partner, the function of machining quality evaluation is integrated. With the function, the simulated result\u2014virtual workpiece\u2014can be \u2018\u2018measured\u2019\u2019 to evaluate the correctness of the NC program, accuracy of tool setting. With the evaluation results, compensation can be carried out to improve the machining quality. In CNC Partner, the values of dimension (including angle), roughness, thread, etc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000016_6.2005-6088-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000016_6.2005-6088-Figure6-1.png",
+        "caption": "Figure 6. Definition of Orbital Coordinate Frame J",
+        "texts": [
+            " }{j are basis vectors of a coordinate frame J , that means an inertia coordinate frame or an orbital frame depending on a situation. The both cases are explained in the following. When J is regarded as an inertia coordinate frame, an origin of the J frame is coincided with vsq , and each axis of the J frame is always equal to the initial V frame represented by 0V . On the other hand, when J is regarded as an orbital coordinate frame, an origin of the J frame is coincided with vsq , and each axis of the J frame is set as shown in Fig 6. 1j , 2j and 3j axis mean a zenith direction, a tangential forward direction and a perpendicular direction to an orbital plane, respectively. In the both cases, coordinate frame K , which basis vectors are }{k , indicates a spin coordinate frame to represent a spin axis direction. The relationship between the coordinate frame J and K is described in the following using a direction cosine matrix. }{}{ jk JKC= (21) where, )()( 21 \u03c8\u03b8 CCC JK = (22) In the Eq.(22), 1C and 2C mean a single axis rotation matrix around 1-axis and 2-axis, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000814_upec.2007.4468943-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000814_upec.2007.4468943-Figure2-1.png",
+        "caption": "Figure 2: (a) stress distribution after modification (b) Prototype spring of the same model in (a)",
+        "texts": [
+            " As can be seen from figure 1, higher stresses are still located almost in the outer side of the inner end of the slit and in the inner side of the outer end making stress concentrations around both sides of each slit. Therefore, the spirals have been modified by curling the inner end of the slit into the centre and outer end out of the centre to give a better solution to these stress concentrations. Many simulations have been done to come up with the best curl of the ends. The curl outwards at the outer ends and inwards at the inner ends are shown in fig 2. Using closing circles of 5mm diameter above the slit in the outer ends and of 3mm diameter under the slit in the inner ends was the best solution and the stress distribution when the spring is at an axial displacement of 8mm, after the curl modification, is shown in figure 2(a). This optimum design has been built as a prototype spring which is shown in figure 2(b). The prototype\u2019s stiffness was tested experimentally and has been compared with FEM results as shown in figure 3. It is clear that experimental results correlate with the simulation, with an acceptable percentage of error, lending confidence to the design methods used. The figure shows that the stiffness at the maximum deformation, 8mm, is 6kN/m. UPEC 2007 - 185 0 1 2 3 4 5 6 7 0 1 2 3 4 5 6 7 8 9 De form ation (m m ) S tif fn es s (k N /m m ) K-e xp K-FE Figure 3: Experimental and FE stress against deformation 5"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003734_j.engfailanal.2012.11.013-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003734_j.engfailanal.2012.11.013-Figure3-1.png",
+        "caption": "Fig. 3. General representation of the mesh used in the modeling with a detailed view of the defect region.",
+        "texts": [
+            " For the modeling the gear was restricted by its inner diameter. The load applied on modeling was derived from the gear operating torque. It was distributed over three teeth adjacent to the flaw and applied in the pressure line direction on a surface of 3670 mm2 over each the three teeth, so that the tooth just above the flaw experienced a stress of 80 MPa, and its neighboring teeth experienced stresses of 65 MPa. A tetragonal quadratic mesh was used in the simulation with a mesh refinement applied at the region of the defect as shown in Fig. 3. To determine the stress intensity factor (SIF) for the crack tip the defect was considered much smaller than the gear (semi-infinite body) and type I loading mode was considered. With this consideration the SIF can be described as a function of applied load (r) and defect dimension (a) according to the following equation: K \u00bc r ffiffiffiffiffiffi pa p \u00f01\u00de FCG tests were conducted in accordance with the ASTM E647 [5] test method using compact tension specimens with thickness (B) of 15 mm and width (W) of 30 mm machined from the core of fractured gear"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000393_9780470061565.hbb032-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000393_9780470061565.hbb032-Figure7-1.png",
+        "caption": "Figure 7. Origin of Lissajous figure.",
+        "texts": [
+            " The excitation signal, expressed as a function of time, has the form E(t) = E0 cos(\u03c9t) (18) E(t) is the potential at time t, E0 is the amplitude of the signal, and \u03c9 is the radial frequency. \u03c9 = 2\u03c0f (19) In a linear system, the response signal, I is shifted in phase and has different amplitude I0: I (t) = I0 cos(\u03c9t \u2212 \u03c6) (20) An expression analogous to Ohm\u2019s law allows calculating the impedance of the system as: I (t) = E0 cos(\u03c9t) I0 cos(\u03c9t \u2212 \u03c6) = Z0 cos(\u03c9t) cos(\u03c9t \u2212 \u03c6) (21) The impedance is therefore expressed in terms of magnitude, Z0, and phase shift, \u03c6. If we plot the applied sinusoidal signal on the x axis of a graph (Figure 7) and the sinusoidal response signal I (t) on the y axis, an oval shape appears known as Lissajous figure. Analysis of Lissajous figures on an oscilloscope screen was the accepted method of impedance measurement prior to the availability of lock-in amplifiers and frequency response analyzers. Using Euler\u2019s relationship exp(j\u03c6) = cos \u03c6 + jsin \u03c6 (22) to express the impedance as a complex function, the potential is described as, E(t) = E0 exp(j\u03c9t) (23) and the current response as, I (t) = I0 exp(j\u03c9t \u2212 j\u03c6) (24) The impedance is then represented as a complex number, Z = E I = Z0 exp(j\u03c6) = Z0(cos \u03c6 + j sin \u03c6) (25) The expression for Z(\u03c9) is composed of a real and an imaginary part"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000113_10934520600689258-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000113_10934520600689258-Figure1-1.png",
+        "caption": "Figure 1: Swarm plate assay for migration experiments in porous media.",
+        "texts": [
+            " After solidification of the agar, the center of the plate was inoculated with 20 \u00b5L of strain KC liquid culture using a micropipette to disperse cells evenly throughout the depth of the agarose suspension. Inoculated plates were anaerobically incubated at room temperature using a GasPak 150 Anaerobic System (VWR Scientific). The typical incubation time required to detect and track the chemotactic migration wave was 96 hours. For migration studies in porous media, the swarm plate assay was modified as shown in Figure 1. A circular screen mesh of height 0.5 inches and radius 2.5 inches was centered in the plate and filled with sterilized Schoolcraft aquifer solids. The packed solids were then saturated with hot agarose solution slowly poured onto the plate. After cooling, the swarm plates were inoculated with strain KC and incubated as described previously. To record images of the chemotactic migration waves, the swarm plates were placed in a transilluminator box (TB) equipped with two 30-cm fluorescent lights (single 8 W, cool white bulbs) to provide diffuse illumination from a 45-degree angle beneath the plates"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003542_j.elstat.2012.04.003-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003542_j.elstat.2012.04.003-Figure11-1.png",
+        "caption": "Fig. 11. Helicopter Model.",
+        "texts": [
+            " Possibly the effect of the finite size of the metal parts of the electric field meter itself may lower the reading in a non-linear and rapidly-changing field. 4.3. Electric field in the presence of a floating-potential helicopter model Since the wing, skid, tail and other parts with small radii of curvature will have a high field close to them; these parts were accurately modeled in the finite-element calculations, in order to obtain reliable results, using the dimensions and layout shown in Fig. 10. The fuselage, and other parts which affect the electric field less, were simplified in the model used for the computation, as shown in Fig. 11a; and similarly in the model built for the experiments (and shown in Fig. 11b). The diameter of the cylindrical cabin was 0.19 m. The measurement of the field in the presence of the helicopter was carried out in the same manner as with the sphere. Again the voltage applied to the line was 200 kV.The measurement line extended from the transmission line in the normal direction and horizontally towards the rotor blades. The electric field distribution was measured and computed as before and the results are shown in Fig. 12 for several different values of d, the distance between the transmission line conductor and the helicopter model\u2019s rotor axle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003530_amr.694-697.503-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003530_amr.694-697.503-Figure2-1.png",
+        "caption": "Fig. 2 Tooth surface modeling of involute Fig. 3 3D modeling of external involute spiral bevel gear spiral bevel gear",
+        "texts": [
+            " The tooth equation of involute curve can be described as follows:                       \u2212\u2212+\u2212\u2206\u2212\u2212+ \u2212+\u2206\u2212\u2212\u2212= \u2212+ \u2212\u2206\u2212+\u2212\u2206\u2212+ \u2212+\u2206\u2212\u2212\u2212 \u2212+\u2206\u2212\u2212\u2212= \u2212+ \u2212\u2206\u2212+\u2212\u2206\u2212\u2212+ \u2212+\u2206\u2212\u2212\u2212 \u2212+\u2206\u2212\u2212\u2212= \u22c5+ \u22c5 + \u22c5 11 cot1 1 1 11 111 cot 11 1 11 11 111 cot 11 1 1 sin)cos()1(cos)sin()1( cos)cos()] 2 1 (sin[ coscos)cos( ]sincos)sin(sin)[cos( sincos)cos()] 2 1 (sin[)1( sin)sin()] 2 1 (sin[)1( cossin)cos( ]sinsin)sin(cos)cos([ sinsin)cos()] 2 1 (sin[)1( cos)sin()] 2 1 (sin[)1( \u03b4\u03d5\u03b4\u03d5\u03b8\u03b8 \u03b4\u03d5\u03b2\u03b8\u03b8\u03d5 \u03b4\u03d5\u03d5 \u03b4\u03d5\u03d5\u03b8\u03b8\u03d5\u03d5\u03b8\u03b8 \u03b4\u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03b4\u03d5\u03d5 \u03b4\u03d5\u03d5\u03b8\u03b8\u03d5\u03d5\u03b8\u03b8 \u03b4\u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03b2\u03b8 \u03b2\u03b8 \u03b2\u03b8 rre srz rr e sr sry rr e sr srx ii j C j Ciij ii CC Cii j Cii j j ii CC Cii j Cii j j \u2213\u2213 \u2213\u2213 \u2213\u2213 \u2213\u2213 \u2213\u2213 (4) The tooth equation of dedendum transition curve can be represented as:                       \u2212\u2212+\u2212\u2206\u2212\u2212+ \u2212+\u2206\u2212\u2212\u2212= \u2212+ \u2212\u2206\u2212+\u2212\u2206\u2212+ \u2212+\u2206\u2212\u2212\u2212 \u2212+\u2206\u2212\u2212\u2212= \u2212+ \u2212\u2206\u2212+\u2212\u2206\u2212\u2212+ \u2212+\u2206\u2212\u2212\u2212 \u2212+\u2206\u2212\u2212\u2212= \u22c5+ \u22c5 + \u22c5 11 cot1 1 2 11 111 cot 11 1 12 11 111 cot 11 1 2 sin)cos()1(cos)sin()1( cos)cos()] 2 1 (sin[ coscos)cos( ]sincos)sin(sin)[cos( sincos)cos()] 2 1 (sin[)1( sin)sin()] 2 1 (sin[)1( cossin)cos( ]sinsin)sin(cos)cos([ sinsin)cos()] 2 1 (sin[)1( cos)sin()] 2 1 (sin[)1( \u03b4\u03d5\u03b4\u03d5\u03b8\u03b8 \u03b4\u03d5\u03b2\u03b8\u03b8\u03d5 \u03b4\u03d5\u03d5 \u03b4\u03d5\u03d5\u03b8\u03b8\u03d5\u03d5\u03b8\u03b8 \u03b4\u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03b4\u03d5\u03d5 \u03b4\u03d5\u03d5\u03b8\u03b8\u03d5\u03d5\u03b8\u03b8 \u03b4\u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03d5\u03d5\u03b2\u03b8\u03b8\u03d5 \u03b2\u03b8 \u03b2\u03b8 \u03b2\u03b8 rre srz rr e sr sry rr e sr srx ii j C j Ciij ii CC Cii j Cii j j ii CC Cii j Cii j j \u2213\u2213 \u2213\u2213 \u2213\u2213 \u2213\u2213 \u2213\u2213 (5) Where, 2,1=j represent external and internal involute spiral bevel gears respectively. Based on the tooth equations of involute curve and dedendum transition curve, the tooth surface modeling of involute spiral bevel gear is drew by using the visualization command in mathematical computation software, as shown in Fig. 2. Then 3D modelings of involute spiral bevel gears have been further completed in SolidWorks system respectively. Fig. 3 shows the three-dimensional modeling of external involute spiral bevel gear. This paper proposed a mathematical model and 3D modeling of involute spiral bevel gears for nutation drive. The equations of basic involute tooth profile have been developed, and then the mathematical models of involute spiral bevel gears have been obtained. The 3D modelings of involute spiral bevel gears have been further completed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002986_s00170-012-4659-1-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002986_s00170-012-4659-1-Figure11-1.png",
+        "caption": "Fig. 11 Schematic of the ECB finishing system for driving tines",
+        "texts": [
+            "1 Finishing process for driving tines As shown in Fig. 1, the central line of the driving tine of the SCMW samples is a cylindrical helix. And the equation of the central line is expressed in Eq. 3 [1]. The section of the driving tine is circle. Suppose the diameter is d: x1 \u00bc m cos t y1 \u00bc m sin t z1 \u00bc nt \u00fe np p t p 2 8< : \u00f03\u00de wherem is the radius of the spiral curve and 2n\u03c0 is the pitch of the spiral curve. The schematic diagram of the ECB finishing system designed for the driving tines is shown in Fig. 11. As shown in Fig. 12, the machining tool has six machining areas which match the driving tines. The section of the machining area is circle. Suppose the diameter is D, and the relationship between D and d is expressed as Eq. 4: D \u00bc d \u00fe 2 h \u00f04\u00de where h is the thickness of the insulating cloth. As shown in Fig. 13, the driving tines could be finished by the aid of a four-axis positioning device. The device consists of an XY stage for positioning a machining tool, a rotating mechanism, and a traversing mechanism for positioning a driving wheel"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure1.5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure1.5-1.png",
+        "caption": "Figure 1.5 Flyweight governor",
+        "texts": [
+            " The pilot\u2019s input is compared with the control surface position (servo actuator output) in a mechanical summing linkage arrangement similar to that shown in Figure 1.4. In a similar manner inputs from an autopilot actuator can be summed with the pilot\u2019s command to provide an auto-stabilization function. The Speed Governor The speed governor goes back more than 200 years to James Watt who invented the \u2018flyball governor\u2019 as a mechanism to control the speed of a steam engine that did not require human intervention. Derivatives of 8 Developing the Foundation this device are used throughout industries using rotating machinery. Figure 1.5 shows such a device in schematic and block diagram form. Today, flyweights rather than balls are more typical offering a more compact design. The basic concept, however, remains unchanged. The rotation from the machinery to be controlled causes the weights to be Block Diagrams 9 thrown outward due to the centrifugal force. This force is compared with an opposing spring force defining the desired operating speed. If the speed exceeds this value the resulting upward motion of the flyweights is used to close the valve supplying the energy to the rotating machinery under control"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000784_acc.2008.4587070-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000784_acc.2008.4587070-Figure1-1.png",
+        "caption": "Fig. 1. The Spatial Configuration of the Helicopter",
+        "texts": [
+            " Finally, the performance of the method is studied through some simulated flight scenarios. The method is also analyzed for associated computational cost of the real-time implementation and a comparison is made with the former analysis of classical gradient descent method. The simplified dynamic equations of a helicopter, which are detailed enough for control development for quasi-steady maneuvers, are introduced. The aerodynamic tractions are assumed to be the control inputs. The spatial configuration of the helicopter is shown in the Fig. 1. Using the Newton-Euler equations of motion, one can link the absolute linear and angular accelerations of the helicopter to the aerodynamic tractions exerted by the main and tail rotors. The inertial position of the helicopter is defined by vector pI , where the index I indicates the vector is expressed in the inertial frame {I}. The equations for the linear acceleration of the helicopter can be formed as the following: p\u0308I = 1 m [RIB(fdB + faB) + fiI ] (1) where RIB is the transformation between the inertial and the body frame B defined by a successive roll (\u03c6), pitch (\u03b8), and yaw (\u03c8) rotations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001902_emobility.2010.5668045-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001902_emobility.2010.5668045-Figure1-1.png",
+        "caption": "Fig. 1. Structure of EV drive system",
+        "texts": [
+            " Based on these demands an exemplary induction machine (IM) is dimensioned and the specific characteristics are presented. A stator-flux-oriented control is introduced that satisifies the difficult challenges and allows very fast torque control in the whole speed range, which is important for safety-relevant features like slip-slide control, electronic stability programs or emergency brake assistance. A modern drive for electric vehicles consists of a threephase two-level inverter feeding a rotating field machine, here an IM (Fig. 1). The machine is connected to the mechanical power train of the vehicle. In contrast to electric drives in traction and industrial use, the inverter is not connected to a constant-voltagecontrolled DC link with a huge capacitor. A battery acts as energy source and is linked by a second-order LC-filter (Lf ,Cf ) to the inverter. The filter reduces the distortion current and avoids thermal overloading of the battery. Due to space restrictions, the filter capacitor is small. Furthermore the internal resistance of the battery source and the ohmic resistance of the wiring cannot be neglected"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001727_0022-2569(69)90006-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001727_0022-2569(69)90006-8-Figure1-1.png",
+        "caption": "Figure 1.",
+        "texts": [
+            " In a second paper [7], which follows we classify the overconstrained R - P - C - P and R - P - P - C mechanisms reported by Dimentberg and Yoslovich [8]. * Departmental Director of Research, Department of Mechanical Engineering, Liverpool Regional College of Technology, Liverpool, England. t Research Assistant at above institution. 261 2 6 2 2. Notation and Displacement Equations for the R - C - C - C Mechanism By applying dual number quaternion algebra, Yang and Freudenstein [1, 2] derived explicit relations to specify the position of the R-C-C-C mechanism for any given rotation of the input link. Figure 1 shows the arrangement of the R-C-C-C mechanism and the sixteen parameters used to specify it. Each pair axis is represented by a unit line vector S~; the common perpendicular between consecutive pair axes ~i and Sj. is the unit line vector hu; the length of the common perpendicular is a u, which is taken as the kinematic link length; the distance between consecutive links a u and aj~is S i measured along the pair axis Si; the angle between consecutive pair axes St and Sj is ~u directed according to the clockwise rotation about au necessary to align ~ with S j; the angle between the consecutive links a~iand ajk is 0~ directed according to the clockwise rotation about ~j"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003610_cdc.2011.6160923-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003610_cdc.2011.6160923-Figure1-1.png",
+        "caption": "Fig. 1: Particle model in a uniform flowfield.",
+        "texts": [
+            " Using Newton\u2019s second law (and assuming unit mass), (5) and (6) can be combined to give s\u0307ke i\u03b3k + skie i\u03b3k \u03b3\u0307k = \u03beke i\u03b3k + \u03bdkskie \u03b3k . Equating the real and imaginary parts yields the equations of motion r\u0307k = ske i\u03b3k s\u0307k = \u03bek (7) \u03b3\u0307k = \u03bdk. Steering and speed control laws are derived for (7). However, in order to implement the control onboard a vehicle, the control laws must be transformed for use in the following dynamics: r\u0307k = vke i\u03b8k + \u03b2 v\u0307k = Tk (8) \u03b8\u0307k = Rk vk = uk, where Tk is the speed control and Rk is the steering control relative to the flow. Using the geometry of the system shown in Figure 1 and the particle models given by (7) and (8), the transformation to express Tk = T (\u03bek, \u03bdk) and Rk = R(\u03bek, \u03bdk) is Rk = \u03bdksk \u2212 sk\u03bek sin(\u03b8k \u2212 \u03b3k) vk + \u03b2 cos \u03b8k cos(\u03b8k \u2212 \u03b3k) + \u03b2 sin \u03b8k sin(\u03b8k \u2212 \u03b3k) vk + \u03b2 cos \u03b8k (9) Tk = sk\u03bek + \u03b2 sin \u03b8kRk vk + \u03b2 cos \u03b8k . (10) With (9) and (10) any control derived in inertial coordinates can be transformed for implementation onboard a vehicle. In this section, we analyze the mapping performance of vehicles traveling in a circular time-splay formation using the metrics defined in Section II-A"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003878_demped.2013.6645747-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003878_demped.2013.6645747-Figure1-1.png",
+        "caption": "Fig. 1. The healthy single cage induction motor.",
+        "texts": [
+            " 4 different cases are investigated: a) The motor is healthy and the voltage supply is symmetrical, b) The motor is healthy and the voltage supply is asymmetrical, c) The motor has a broken bar and the voltage supply is symmetrical and d) The motor has a broken bar and the voltage supply is asymmetrical. When the voltage supply is symmetrical the applied voltage is 380V for all phases. In the asymmetrical supply condition the first and second phases have 380V, while the third has 353.5V. The simulated, healthy, single cage induction motor is presented in Fig. 1. In this section two cases are studied. The single cage induction motor under healthy and with one broken bar operation at 1460rpm. In Fig. 2-a, the FFT of the line current is presented for both cases. The broken bar fault diagnosis offers reliable signatures close to the stator MMF harmonics at 50Hz, 250Hz and 350Hz, totally agreeing with previous knowledge. Best fault signature can be considered the one at 47.6Hz with -31.4dB amplitude (Fig.2-b). a) In Fig. 3-a, the zero-sequence current spectrum for the two cases is presented"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000152_piee.1973.0142-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000152_piee.1973.0142-Figure2-1.png",
+        "caption": "Fig. 2",
+        "texts": [
+            " If the rotor speeds are held constant and the saturation is neglected, the analysis will be called 'linear transient analysis', since the system equations are linear and time varying. Linear transient analysis is used in fault studies. Practically every study concerning synchronous machines assumes only fundamental-frequency m.m.f. distribution in the airgap. This assumption will be made in the following analysis. Magnetic saturation is also ignored here. The analysis of machines with m.m.f. harmonics is indicated in Section 5. 3.1 Machine equations in phase parameters The equations given here apply to a synchronous machine shown in Fig. 2. The rotor circuits on the d axis represent the field and damper windings. Either two damper windings or a quadrature-axis field winding and a single damper winding are represented by the rotor circuits on the q axis. Actually, there is no restriction on the number of rotor circuits that can be considered, but two circuits on each axis are considered to be adequate. Synchronous machine 649 The machine parameters Ls, Lr ,M,Rs and Rr are now defined: cos 20r cos (20r - \u2014\\ cos (20r + \u2014\\ M = ls - Laao Labo L a b 0 Labo L a a o La^o + L Labo Labo L a a o Lf Lfh 0 0 Lfh Lh 0 0 0 0 Lg Lgk Lgk W 'aa2 cos cos \\ r + T",
+            " 16 and 11 are transformed into di ' dt where die - RrLf ^ = vr (37) (38) Since Lg and Lg\"1 are cyclic symmetric matrices, CjLs'C1 and C\\Lg\"1C1 will turn out to be the diagonal matrices As' and As'~i defined by L\" 0 0 0 L\" 0 0 0 h'A where L\" = La'a - L^b, Lg = written as Is = = c1is where Is is defined by Is = Ag-i(M'Lfty and M' = (39) + 2L^b. Is can be re- (40) (41) (42) Substituting eqn.40 in eqns. 37 and 38 gives, finally, dt dt = v s s dt Note that eqns. 39,41, 43 and 44 apply to a machine where the stator phase windings are replaced by a, (i and 0 windings. 652 The mutual inductances between these windings and the rotor windings are given by the matrix M'. For the synchronous machine of Fig. 2, we obtain where (45) The electromagnetic torque is given by \" COS \u2014sin - 0 h + Iq -sin ( cos -0 9r~ 9r where dl - sin ( \u2014cos _ 0 + Iq - cos \u2014sin _ 0 9r~ h (46) (47) Expanding eqn.46,we obtain Te = -L\"{(ia cos 6r - ip sin 0r)Iq - (ia sin c o s (48) The transformed circuit model for a/30 components is similar to that shown in Fig. 1. Note that the current source in the zero-sequence network is zero. Further, the zero-sequence component of the armature current i0 has no influence on the rotor flux linkages i//r, since the mutual inductances between the zero-sequence coil and the rotor coils are zero",
+            " 6, JUNE 1973 cos 20, 2 0 r - \u2014 \\ cos (20r +\u2014) cos 20,r 3 / \\ r 3 / J cos (20r + \u2014J cos 20T3 cos (55) 653 The matrix Lg,<when diagonalised,has the-elements given by eqn. 39. By comparing it with eqn. 57, it can be seen that' L\" = L^. Further, by defining \u2022 machine is shown in Fig. 3B. For this circuit, the state equations can be expressed from the standard networkanalysis approach. These are: cos (er-\u2014\\ cos (eT + \u2014 3/ \\ r 3 s1 = J- sin sin 0V \u2014 the current-source vector Is can be expressed as dt (63) Is = Idc + Iqs (58) where 1$ and Iq are defined in eqn. 45. The expressions for 1\u0302 and Iq can be rewritten in terms of the dq parameters. Referring to Fig. 2, the mutual inductances between coils in the d and q axes are expressed as Mdf=(A/f)Maf Mdh=(V /f)Ma h Mq g = [ , / - Mnt = U / -xqk The rotor-coil equations can be rewritten in terms of the dq parameters and the newly defined variables \u2022 (59) It is to be noted that the machine equations are generally written in per-unit form. There is an additional convention that is commonly used with regard to field voltage. The per-unit field voltage is defined by the equation Rf vbase *df vbase per unit where vD a s e is the base voltage. Summarising, the final equations for the machine shown in Fig.2 are dt c4i//k (60) dt b 3 i d b4vg + b 5 i q dt a8i//k + b 6 i q (61) Te = - x \" ^ - iq^) d6 . d6 wn \u2014 = 6 \u2014 =\u2014-(-D6 + T m - T p ) dt dt 2H m e (62) The coefficients a, b,c and \u00a3 are defined in Appendix 9.1. 4 . 1 TRANSIENT ANALYSIS Formulation The machine model derived in the previous Section is utilised here to show the ease of formulation of the state equations in transient analysis. Consider the system shown in Fig. 3A. The equivalent circuit of the stationary network including the circuit model for the stator coils of the 654 Eqns"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001022_j.jsv.2008.05.023-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001022_j.jsv.2008.05.023-Figure5-1.png",
+        "caption": "Fig. 5. Nonlinear spring stiffness function of the bearing.",
+        "texts": [
+            " A gear element, which is a function of the relationship between the relative displacement S, impact force fi (Fig. 3), and damping force fd, has been developed. The dynamic behavior of the bearings is described by the contact force fB and damping force fc, which comprise three dofs (axial and lateral direction): fB \u00bc kB\u00f0e\u00dee (13) where e is the relative displacement between the shaft and the housing. kB is a spring stiffness function, which is derived from the loading\u2013displacement experiment. The nonlinear spring characteristic in the axial direction is shown in Fig. 5a where, fBx denotes the contact force and ex denotes the relative displacement at the axis of the bearing. In the numerical analysis, the spring function kx in the axial direction is described by the approximating the linear spring coefficients k1 and k2 (Fig. 5b): kx\u00f0ex\u00de \u00bc k1 0pjexjpa k2 jexj4a ( (14) Based on the experiment, the spring stiffness function is approximated as linear in the lateral direction: fc \u00bc CB _e (15) where _e is the velocity of e and CB is the coefficient of damping. Shafts are modeled using beam element (Fig. 6). Since the torsion stiffness of the shafts in the x direction is high, a small number of motion nodes are enough to express the torsional motion. Thus, a large number of nodes on shafts rotating in x direction are not necessary as in the axial and lateral directions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002693_iecon.2010.5675183-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002693_iecon.2010.5675183-Figure1-1.png",
+        "caption": "Fig. 1. Quanser Helicopter of two DOF",
+        "texts": [
+            " In the following paper a design of robust controller using DK iteration is presented, modeling parametric and structured uncertainty. The original model of the system is a nonlinear one, and then it is linearized around an equilibrium point. In II the mathematical model of the system is presented, In III the controller by DK iteration is commented, in IV the uncertainties of the system are shown, in V the simulations of the system with the controller are presented and finally, the conclusions are shown in VI. In Figure 1 the Quanser helicopter of two DOF is shown. It is mounted on a fixed base with two propelled that are driven by DC motors. The front propeller controls the elevation of the nose over the pitch axis. The tail propeller guides the rotation motion around the yaw axis. The motors correspond to each one of the actuators of the propellers. The pitch and yaw motors voltage are \u00b124V and \u00b115V respectively. Some others specifications like masses, torques, inertias, and others, are presented in [6]. In order to obtain the mathematical model of the system, A 978-1-4244-5226-2/10/$26"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001319_j.ijpvp.2007.03.004-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001319_j.ijpvp.2007.03.004-Figure3-1.png",
+        "caption": "Fig. 3. Displacement pattern of the flange geometry 1. (a) Displacement of (b) Characterization of the displacement of geometry 1.",
+        "texts": [
+            " Joshi et al. / International Journal of Pressure Vessels and Piping 84 (2007) 405\u2013411 407 node. Now if the bolt ends are rigidly restrained the temperature difference required to generate a force of F is given by DT \u00bc F EAa . Here E is Young\u2019s modulus of the material, A is the cross sectional area of the beam and a is the coefficient of thermal expansion. Due to the compliance of the flange, DT needs to be adjusted to achieve the desired F. We have taken two different sets of bolt geometries. Fig. 3 (a) shows the displacement of the flange for an internal pressure of 10MPa. The total bolt area was taken as 0:0628m2 with a net preload of 1:3183 107 N. As seen from Fig. 3 (a), the rectangular part of the flange cross section rotates like a rigid body. Thus, for the simple calculation we will present for geometry 1, the flange displacement merely involves an angle y and a separation x (see Fig. 3 (b)). The overall problem is nonlinear primarily because of making or breaking mechanical contact. Linear elasticity does not distinguish between deformed and undeformed configurations, and cannot directly handle contact. Moreover, the present problem has a significant nonzero bolt preload. Under a change in configuration, the preload gives terms that are formally of the same order of magnitude as other terms in the equilibrium equation. However, if overall displacements and rotations are small, and if stresses everywhere are within elastic limits, then the problem still retains significant aspects of linear elasticity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure9-1.png",
+        "caption": "Figure 9. Statico-dynamic balancing with three identical rotating masses.",
+        "texts": [
+            " L e t n ' = 2 ; t h e n k = l . the simultaneous solution of the equations of type (4), (5) and (7), the second trajectory becomes fully d e t e r m i n e d . The above indicates that, in the statico-dynamic balancing of a plane mechanism with three counterweights, it is possible to select the trajectories of all three counterweights. We shall balance a mechanism with three rotating masses, assuming that i i t i and that their centers move along the circumference of the circle, described about the center U~. with radius p (Fig. 9). First of all, let us find the position of the center of mass (B~) of the counterweights for the ith position of the mechanism. It is obvious that S :--m~ ]1. 3m' ' : i;:)::: ~ m ' Figure 11). The determination of the center of mass of three masses. Point B is the center of mass of the three identical counterweights, distributed at points vertices of the triangle F and A, coincides with point H- - t he center of the side FA. The center B of all three counterweights lies on the straight line DH. Since the mass of one counterweight m' is concentrated at point D, and H is the center of mass of two such counterweights, then the following ratio of lengths will hold true: BD 2 HB 1 Consequently, the center of mass B of the three identical counterweights is situated on the median of the triangle at a distance of 3 z the median from the vertex. Point B is the point of intersection of all three medians. Now return to the mechanism to be balanced (Fig. 9). Select arbitrarily the position of one of the counterweights, for example, the point Di, and connect points D~ and B i with a straight line. Di is the center of one of the counterweights, and Bi is the center of mass of all the counterweights. Thus, the segment DiB~ comprises ] of the median. Now find the magnitude HiBi - BiDi 2 and determine the point H i, which is the center of that side of the triangle, on the ends of which the counterweights are located. Their centers, however, move in a circle described about the point O(U~y) with radius/9"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002691_icca.2010.5524420-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002691_icca.2010.5524420-Figure5-1.png",
+        "caption": "Fig. 5 Mathematical model of the 3-rotor aircraft",
+        "texts": [
+            " The authors are with National Key Laboratory of Science and Technology on Holistic Control, Department of Automatic Control, Beijing University of Aeronautics and Astronautics, Beijing 100191, China. (Email: fphbuaa@126.com, wangxinhua04@gmail.com, kycai@buaa.edu.cn) As it is shown in Fig. 2, the aircraft has three rotors and two commutating devices which are respectively driven by motors and small actuators. From components 4-8, it is shown that the three ro-tors rotate synchronously with the same angle \u03b2 (shown in Fig. 5). Firstly, actuator 4 drives the driving gear 5, then, the driving gear 5 drives the transmission shaft 6, finally, the transmission shaft 6 drives the front driven gear 8 and the tail driven gear 7. So the three rotors can rotate synchronously with the same angle \u03b2 . The main components are explained in Table 1. In order to obtain the dynamic model of the tri-rotor aircraft, we need the values of the aerodynamic parameters needed in the flight simulation near the real condition. We use software tool Fluent to get the values",
+            " 4) of the simulation targets could not come to a pre-seted value such as 10\u22125, then, we need to turn back to the first step, and make the mesh denser, repeat the second step until the residuals of the simulation targets come to the pre-seted value. Third: Postprocessing By using the postprocessing tools in Fluent, we get the values of aerodynamic parameters which are presented as follows. The lift and drag forces of selected wing: L = 0.5CLS\u03c1 ( x\u03072 b + z\u03072 b ) , CL = CL0 +CL\u03b1 \u03b1, CL0 = 0.3137,CL\u03b1 = 0.7025, D = 0.5CDS\u03c1 ( x\u03072 b + z\u03072 b ) , CD = CD0 +AC2 L, CD0 = 0.00182,A = 0.1587, (1) where \u03b1 is the attack angle of the wing(see Fig. 5), S is the area of wing, \u03c1 is the density of air, L and D are lift force and drag force respectively. x\u03072 b , z\u03072 b can be seen in Fig. 5. The lift force of rotors: fi = \u03ba\u03c92 i ,(i = 1,2,3) ,\u03ba = 2.051\u00d710\u22125, (2) where \u03c9i is the angular velocities of the rotors. The moments of rotors: Qi = \u03b3\u03c92 i ,(i = 1,2,3) ,\u03b3 = 1.225\u00d710\u22127. (3) The drag force of commutating device: fri = \u03ba1 ( \u03b40 + c\u03b4 2 ) \u03c92 i ,(i = 1,2) , \u03ba1 = 1.2675\u00d710\u22126,\u03b40 = 0.238,c = 1.923, (4) where \u03b4 is the deflection angle of recti-block. The moment of commutating device: Qri = \u03b31\u03b4\u03c92 i ,(i = 1,2) ,\u03b31 = 4.5375\u00d710\u22127. (5) IV. MODELING OF THE AIRCRAFT As shown in Fig. 5, ogxgygzg is the inertial coordination, and obxbybzb is the airframe coordination. ob is the center of the airframe, m is the mass of the aircraft. p = (xg,yg,zg) T is the projection of airframe center in the inertia coordinate. Then we have the dynamic and kinematic model as follow:    p\u0307 = v p\u0308 = ge3 +RF/m R\u0307 = RS(\u2126) \u03c4 = J\u2126\u0307+\u2126\u00d7 J\u2126 (6) where S(l) is the associated skew-symmetric matrix of vector l such that S(l)\u2126 = l \u00d7\u2126. v is the projection of the airframe velocity in the inertia coordinate, \u2126 is the projection of the airframe angular velocity in the airframe coordination"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001465_tmag.2007.915810-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001465_tmag.2007.915810-Figure4-1.png",
+        "caption": "Fig. 4. Measured and calculated BH hysteresis loops in the transverse direction.",
+        "texts": [],
+        "surrounding_texts": [
+            "In magnetostatics, Maxwell\u2019s equations to be solved on a domain are (6) For the sake of simplicity, the conditions on the boundary of domain are given by . The behavior law can be written such as . In non-ferromagnetic material, this relation becomes whereas in a ferromagnetic material, the vectorized JA model is used. To solve the problem, the vector potential formulation is used [6]. By applying the Galerkin method, the weak formulation is (7) where is the magnetic vector potential introduced using is a test function, and the current density. Let us consider now a tetrahedral mesh of domain D. If the magnetic vector potential is discretized in the edge element space, we obtain (8) with the interpolation function associated to the edge and the number of edges not located on a boundary. Actually, due to the boundary condition, the tangential component of is equal to zero on . Then, by using the interpolation functions as test functions, equation (7) can be written under the form (9) To solve this nonlinear problem, the NR method is applied [6]. It leads to the following nonlinear equation system (10) (13) The system is solved until the error convergence criterion for is reached. The coupling with the circuit equation can be done easily by an equation linking the linkage flux flowing the winding and the supply voltage [9]."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001086_s11044-007-9053-7-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001086_s11044-007-9053-7-Figure2-1.png",
+        "caption": "Fig. 2 The lengths ai and nodes of the Andrews\u2019 system",
+        "texts": [
+            " It is a bit more surprising that there is another subsystem producing singularities: the one formed by the nodes 3, 6, and 7. Define h1 := \u2212p5 + p1 = a6(c6c7 \u2212 s6s7) + a7s7 \u2212 a3s3 + w1 \u2212 b1, h2 := \u2212p6 + p2 = a6(s6c7 + c6s7) \u2212 a7c7 + a3c3 + w2 \u2212 b2, h3 := p9 = c2 3 + s2 3 \u2212 1, h4 := p12 = c2 6 + s2 6 \u2212 1, h5 := p13 = c2 7 + s2 7 \u2212 1. It is important to note that h1, h2 contain only angles 3, 6, and 7, and therefore only p9, p12, p13 are relevant to them. As parameters we now have not only the lengths a3, a6, a7, but also b1, . . . ,w2, i.e., the positions of the fixed nodes A and B in Fig. 2. Let J367 be the ideal generated by h1, . . . , h5. We will proceed in a similar way as with the subsystem J4567. First we will consider the singularities of the subsystem J367 using the following product order: J367 \u222a FJ367 \u2282 Q [ (c3, s3, c6, s6, c7, s7), (a3, a6, a7, b1, b2,w1,w2) ] . (15) The relevant Gr\u00f6bner basis G contains 96 generators of which two are especially interesting: g12 = c6a6a7, g1 = 4\u220f i=1 zi where z1 = (a3 \u2212 a6 + a7) 2 \u2212 |b \u2212 w|2, z2 = (a3 + a6 + a7) 2 \u2212 |b \u2212 w|2, z3 = (a3 + a6 \u2212 a7) 2 \u2212 |b \u2212 w|2, z4 = (a3 \u2212 a6 \u2212 a7) 2 \u2212 |b \u2212 w|2",
+            " Although it is not immediately apparent, in that situation there also arises a new condition among the parameters: a3 = |b \u2212w|, i.e., \u201ca3 equals the distance between A and B\u201d. Note that here the Fitting ideal FJ4567 has not been used at all, contrary to the T5 calculations. I3 corresponds to y6 = y4 + \u03c0 and y5 = y7 \u2212 \u03c0/2 so that now nodes P3 and P4 coincide. Then again, I1 corresponds to y6 = \u2212y4 and y5 = y6 + y7, which interestingly is not a singularity but merely expressing a symmetry in the system due to a4 = a5 = a6 = a7. 4.6 Other subsystems Now contemplating Fig. 2 we see that it would be possible to find other singularities by analysing still other subsystems. For example, the subsystem corresponding to rods 3, 4 and 5 is by symmetry similar to subsystem 367: we simply exchange the roles of variables and parameters associated to rods 4 and 6, and 5 and 7. Furthermore, we could consider other subsystems formed from different \u201cpaths\u201d between the nodes A,B,O: i.e., subsystems J123, J1245, J1267. Again, by symmetry the system J1267 is completely similar to J1245, but cases J123 and J1245 give new singularities"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001104_09544062jmes817-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001104_09544062jmes817-Figure5-1.png",
+        "caption": "Fig. 5 All forces and moments acting on a ball",
+        "texts": [
+            " Using the Newmark-\u03b2 method, the differential equations of motion can be solved, and the transient responses at every time increment are obtained. Because the displacement of a bearing is equal to the displacement of a rotor, the dynamic characteristics of the ball bearing are impacted by vibration of the rotor system. If the vibration displacements of the rotor system are obtained, dynamic properties of the bearing can be computed, considering the vibration of the rotor system. The contact deformation and contact angle between balls and races are computed using the above formula. Figure 5 gives the forces acting on a ball, where Mxi, Myi, and Mzi are the inertia moments of the ball, Fdi is the fluid drag force, and T1i and T2i the traction forces. Fdi can be obtained from reference [16]. An accurate traction model is used, in which non-Newtonian fluid characteristic is considered, and the traction equation is given by [17] \u03bc = (A + Bs)e\u2212Cs + D (26) where s is the spin-to-roll ratio of balls; coefficients A, B, C , and D are determined for a wide range of lubricants from experiment data"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000444_j100664a022-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000444_j100664a022-Figure1-1.png",
+        "caption": "Figure 1. Electrodesolution interphase. The adsorbed anion i s shown with two adsorbed primary waters of hydration; other P W ~ O U ~ ~ ~ N I ~ water is not shown.",
+        "texts": [
+            " 76, No, 20, 1972 three ion-metal interactions have the greatest exponential depenclenco on distance of all the contributions. The first of two modifications in this paper to the theory of Antlcroen and Bockris occurs in the determination of ~~krc equilibrium separation distance R,. Instead of i i ~ ing only the ion-metal interactions to determine Re, :)I1 contributions to AG\" were included. 'This effect is most evident in the case of F- adsorption. At the point whew the ion-metal intcractions are a minimum, the loss in hydration for this ion is about 25%. I n other. words, Lhe angle (b in Figure 1 must be large for. t>hat vdxe (jf R, if R,, the separation distance of\" the water nj?oIecule from the electrode, i s to remain near its equilibrium value. 'The result is a loss in energy of about 1.8 kcal/mol; that this is large can be seen by obserk Ing thst for C1-- the change in energy is O ~ Y eboul, 6 k cal/r-~~d tvheii ion--metal iiiteractions are a t a ~ ~ n i ~ u ~ . ' h e 1 rge loss in energy for F- is mough to make A,6\" i v ~ and there is MI atlsorptjoii ad Lhsl, dis1,ance. I er, when the ion i s further from the metal, Sewer walers of hydration must be removed frcm the hydrated ion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003654_iros.2012.6385631-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003654_iros.2012.6385631-Figure3-1.png",
+        "caption": "Fig. 3. Ground reaction force working on support leg: a Humans\u2019 walking: The time history of the ground reaction force has the two peaks. First peak is created by the force of impact between the foot and the ground, and second peak is created by pushing the ground. b and c Walking controlled by tacit learning: The walking gait discussed in [7] did not push the ground by support leg. Thus, one peak appeared in b. The walking gait in this paper can push the ground at almost the same timing of the humans\u2019 walking gait described in c.",
+        "texts": [
+            " (12) to (15) though the controller use only the environmental signals. This feature implies that the equilibrium state is automatically tuned when the model is changed. This ability shows the advantage of tacit learning over conventional model-based approaches when we deal with a plat where modeling errors should be a critical problem for a stable control. Let us consider the bipedal walking gait that pushes the ground by a toe of a support leg. This walking gait is the typical way for human being as shown in Fig. 3 a in which the ground reaction force working on the support leg is illustrated. The reaction force in vertical direction in Fig. 3 a has the two peaks, first peak is appeared by the force of impact between the foot and the ground at the moment of landing, and second peak is the force pushing the ground by the toe at the moment of losing a contact with the ground. This class of walking gait easily makes falls of a body because the point of effort of the reaction force is far from the center of gravity (COG) of the body and a torque around COG is created in the case where the reaction force is not pointed in the direction of COG. Thus, the modeling errors or the model changes should be a critical problem for stable walking. The walking gait learned in [7] was the gentle one where the support leg was carried up without pushing the ground. The ground reaction force of this walking gait has only one peak as illustrated in Fig. 3 b. We apply tacit learning to create the walking gait of pushing the ground by the toe of the support leg to discuss the ability of tacit learning that can estimate the robot model through the behaviors. We used the 36DOF humanoid robot shown in Fig. 4 in the experiments. To learn the bipedal walking gait pushing the ground, we use the same method discussed in [7] where the joints of the swing leg were specified and the joints of support leg were unspecified dividing the walking step into the 8 phases",
+            " The experimental movies can be seen at [18]. Fig. 6 describes the time history of the right toe angle. The toe was moved randomly at the beginning of the experiment because the motion of swing leg disrupted the standing balance of the robot and made the robot fall. The motion of the toe joint gradually became the rhythmical one as illustrated in Fig. 6 b and eventually the periodic motion appeared after the robot learned the walking. The motions of all other joints showed the similar process of learning the motions. Fig. 3 c illustrates the ground reaction force in this experiment after the robot learned the bipedal walking. We can say from the shape of the graph that the support leg pushed the ground at the moment of losing the contact between the toe and the ground. The experimental result that kept walking by pushing the ground suggested that the forces by pushing the ground were mainly used the translational movement of COG and the torque that rotate COG scarcely generated. All living organisms can learn the appropriate behaviors adapted to the environment through their daily activities"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure7-1.png",
+        "caption": "Fig. 7 Tooth surfaces of the longitu by dual cutter heads",
+        "texts": [
+            " The basic design parameters of the xample gear set are listed in Table 1, while the calculated mahine settings for the gear generator in Table 2. The ratio of the oll angle for the gear and pinion is =0.259, while the initial 31008-6 / Vol. 132, MARCH 2010 om: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/28/201 angles of the cutter blades are l=\u22122 deg and r=182 deg for the left and right cutters. The generated longitudinal cycloidal helical gear set obtained by the proposed mathematical model is illustrated in Fig. 7 a , which also shows that simply adjusting the machine settings of the gear generator enables the production of either a concave or convex tooth trace. The crowning quantity at the midsection of the tooth flank is represented in Fig. 7 b , which clearly demonstrates that the tooth profile modifications are com- inal cycloidal helical gear generator al cycloidal helical gears produced itud din Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use p v t e h i b c = c a a = d r = t O c i b l s 1 s a t m g s p i g J Downloaded Fr arable to the tooth flank of an involute helical gear. Example 2. This example applies the TCA technique under arious meshing conditions to explore the kinematic characterisics of the proposed double-crowned gear set"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002181_te.2010.2043844-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002181_te.2010.2043844-Figure4-1.png",
+        "caption": "Fig. 4. Transformer-based -source driving a Y-load.",
+        "texts": [
+            " The procedural introduction, distributed with the experiment, included sections reviewing simple three-phase systems, balanced loads, equivalence, and synchronous motor basics. Since three different types of three-phase source were being used, a supplemental description of the appropriate source and its operation was placed at each lab station. Students were encouraged to be patient with the instructors and the prototype sources and to ask for help in understanding and operating each source. Students were first directed to connect a balanced Y-load consisting of a resistor and a F capacitor in series for each phase (shown in Fig. 4 with the two-transformer source) and to operate at 60 Hz. The line and phase voltages and currents at the load were determined using both a multimeter and an oscilloscope. Magnitude and phase comparisons were made. Total complex power delivered to each phase of the load was computed. The phase sequence was determined then reversed so that comparisons could be made. In order to study -connected loads, students were asked to configure their source as a positive sequence source. Students determined the -connected load equivalent ( F in series with ) to the previous balanced Y-connected load and connected this to their source (shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002323_978-3-642-25486-4_25-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002323_978-3-642-25486-4_25-Figure1-1.png",
+        "caption": "Fig. 1. Principal structure of the DELTA robot with vector loop visualization (left side). Definition of local reference frames for each kinematic chain (right side) [6].",
+        "texts": [],
+        "surrounding_texts": [
+            "The calculation of the inverse kinematics starts with the vector loop equation for each kinematic chain: .= + + \u22121i 2i i ip l l a b (1) We assume that all joint coordinates of the base frame as well as all joint coordinates of the end-effector lie on circles with radii a and b but have the same angular allocations. In this case, transformation matrices from the global reference frame into the leg-sided reference frames can be introduced [5]: cos sin 0 sin cos 0 . 0 0 1 \u03b1 \u03b1 \u03b1 \u03b1    = \u2212      i i i i i 0T (2) At this point a new auxiliary vector can be calculated: 0 . 0 \u2212 +   = \u2212 + + = + \u22c5      a b i i i i i i i i 0r a p b T p (3) During the inverse kinematic calculation the end-effector position vector p is known. The vector can also be expressed with: ( ) ( ) ( ) ( ) ( ) ( ) ( ) cos sin cos 0 cos . sin sin sin \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5    \u22c5 \u22c5 \u22c5 +    = + = + \u22c5       \u22c5 \u22c5 \u22c5 +    1 1i 2 3i 1i 2i 2 3i 1 1i 2 3i 1i 2i l l l l l i i i i 1i 2ir l l (4) While evaluating the second component equations (3) and (4) lead to the equation for angle : 1 sin cos cos . \u03b1 \u03b1\u03d5 \u2212  \u2212 \u22c5 + \u22c5 =     i i 3i 2 x y l (5) The second auxiliary variable can now be computed with the help of the magnitude of vector : ( ) 2 2 2 2 2 1cos . 2 sin \u03d5 \u03d5 \u2212  + + \u2212 \u2212  =  \u22c5 \u22c5 \u22c5  i i i i,x i,y i,z 1 2 2i 1 2 3i r r r l l l l (6) The actuation angles , which are the final results of the inverse kinematic calculations are also resulting from the evaluation of equations (3) and (4): ( ) sin cos sin sin tan . sin sin sin cos \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u2212 \u22c5 \u2212 \u22c5 \u22c5 \u22c5 + \u22c5 \u22c5 \u22c5 = \u2212 \u22c5 + \u22c5 \u22c5 \u22c5 + \u22c5 \u22c5 \u22c5 i i i 1 i,z 2 3i 2i i,z 2 3i 2i i,x 1i i i i 1 i,x 2 3i 2i i,z 2 3i 2i i,x l r l r l r l r l r l r (7)"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure23-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure23-1.png",
+        "caption": "Fig. 23. Riveted joint deformations after upsetting",
+        "texts": [
+            " Using this procedure, the iteration process is done simultaneously to satisfy both the contact constraints and global equilibrium using the Newton-Raphson method. This procedure is accurate and stable but may require additional iterations. Numerical analysis The numerical calculations were performed for two cases of upsetting (w1, w2) distinguished by the height of the formed rivet head (Table I, Fig. 21). Methods for FEM analysis of riveted joints of thin walled aircraft structures 959 Riveted joint deformations are shown in Fig. 23, while rivet deformations for two cases of upsetting are given in Fig. 24. In the first case (w1), the height of formed rivet head has maximum value, when the diameter reaches nominal value [17]. For the case w2, diameter of formed rivet head reaches maximum permissible value according to riveting process manual [2]. Rivet upsetting causes metal sheet joining and filling of the rivet hole. Because of increasing fatigue life of the riveted joint, the rivet upsetting and compressive stresses distribution in the sheets around the hole should be uniform and exceed yield stress level [18,19]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure5-1.png",
+        "caption": "Figure 5. Arc-driving path, when l1<l2.",
+        "texts": [
+            " The radius value, R1, can be determined by finding a circle that passes through the target point T and at the same time is tangent to the target direction line and also to the current direction line, which is the Yw axis. R1 is calculated as follows; (1) Since, PA=PT=TB/sin(\u03b8)=|a|/sin(\u03b8) and R1=CA=PA tan(\u03c8)=|a|/sin(\u03b8)\u00d7tan(\u03c8) where, (2) The y-coordinate position of point A(0, Ay) (which is on the Yw axis) at which the arc path starts is given by Ay=Py\u2212PA=Py\u2212|a|/sin(\u03b8) (3) Case 1-2 (Py \u2265 0 and |\u03b8| \u2264 90o and l1 < l2) Figure 5 shows the case of l1<l2. For a vehicle to travel from the current position C(0,0,0) to the target position T(a,b,\u03b8), it should follow an arc path of radius R2 from the current position C(0,0,0) until point B, from which the vehicle should follow a straight line (equivalently an arc of infinite radius) up to the target position T(a,b,\u03b8). The radius of the arc, R2, is calculated as follows; (4) where, l1=Py The coordinate position of point B (denoted as B(bx, by)) at which a straight travel path starts, is calculated as R1= a sin\u03b8 ---------- tan 180o \u03b8\u2013 2 -------------------\u239d \u23a0 \u239b \u239e \u03c8 = 180 \u03b8\u2013( ) 2 ---------------------- R2= l1( )tan 180o \u03b8\u2013 2 -------------------\u239d \u23a0 \u239b \u239e follows: PC=PB=l1=Py Bx=PD=PB \u00d7 cos(90\u2212\u03b8)=l1 cos(90\u2212\u03b8) By=l1+BD=l1+l1 sin(90\u2212\u03b8) (5) Case 2-1 (Py<0 and \u03b8 >0) If Py<0, a path of one arc is not sufficient to get to the target position and meet the target direction"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000966_la802967p-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000966_la802967p-Figure8-1.png",
+        "caption": "Figure 8. Explanation of the parametrization of the surface that is used in order to calculate the elastic deformation energy for the prescribed director field. The director field in a point (x\u0303, y\u0303) in the bulk lies on the circle section that intersects the profile y(x) perpendicularly in (x0, y0). This circle has center (0, ) and radius r. \u03c6 is the angle that the line connecting a point on the circle section and (0, ) makes with the horizontal. y\u2032(R) ) 0 is presumed,giving ) 0 and \u03c6 \u2208 (0, \u03c0/2) for points on the bottom circle section.",
+        "texts": [
+            " In the strong-anchoring limit, we for simplicity but without loss of generality presume symmetric homeotropic anchoring because it allows for a convenient mathematical construction of the director field, which is worked out in this Appendix. Because of the invoked equal-constant approximation, the result also applies to planar anchoring. From the profile equation, we next derive interesting properties such as the decay length and the contact angle. (i) Parametrization For a point (x\u0303, y\u0303) in the bulk of the rise region, we find the section of a circle that has its center on the y-axis, and intersects the profile y(x) perpendicularly in (x0, y0); see Figure 8. The line tangent to this point intersects the vertical axis at height ) y0 - y\u2032(x0)x0. The profile is assumed to decay completely within a finite distance R from the wall. We formally let Rf \u221e at the end of the calculation. The points (x\u0303, y\u0303) satisfying x\u03032 + (y\u0303 - )2 ) x0 2 + (y0 - )2 ) x0 2(1 + (y\u2032(x0))2) are on the same distance from (0, ), so they are on the same field line. We then find for the director field at (x\u0303, y\u0303): n(x\u0303, y\u0303)) (- y\u0303- \u221ax\u03032 + (y\u0303- )2 , x\u0303 \u221ax\u03032 + (y\u0303- )2) (16) where depends on x\u0302 and y\u0302 implicitly",
+            " The angle that the tangent makes with the horizontal is \u03c6. We now shift from the coordinates x\u0303 and y\u0303 to coordinates x and \u03c6. These are not the usual polar coordinates because the origin is not fixed. Abbreviating y\u2032 for y\u2032(x), we find x\u0303) r cos \u03c6) x\u221a1+ y\u20322 cos \u03c6 (17) y\u0303) - r sin \u03c6) y- y\u2032x- x\u221a1+ y\u20322 sin \u03c6 (18) The Jacobian of our coordinate transformation is |\u2202(x\u0303, y\u0303) \u2202(x, \u03c6) |) x(1+ y\u20322)+ x2y\u2032\u2032(y\u2032+ \u221a1+ y\u20322 sin \u03c6) (19) The boundaries for x are (0, R), the upper bound for \u03c6 is \u03c6 ) \u03c0/2; the lower bound follows from Figure 8: tan \u03c6 ) ( - y)/x ) -y\u2032, so \u03c6 ) -arctan y\u2032. (ii) Elastic Free Energy The director field can easily be expressed in terms of the new variables, which gives for the elastic deformation density (66) Sokalski, K.; Ruijgrok, T. W. Phys. A 1982, 113, 126\u2013132. (67) Allen, M. P. unpublished results. (68) Hariharan, P. Basics of interferometry; Academic Press: London, 1992. (69) Caber, P. J. Appl. Opt. 1993, 32, 3438\u20133441. (70) Harasaki, A.; Schmit, J.; Wyant, J. C. Appl. Opt. 2000, 39, 2107\u20132115. fE ) (\u2207 \u00b7 n)2 + (\u2207 \u00d7 n)2 ) 1 x2(1+ y\u20322) (20) fE diverges for xf 0, so we introduce a cutoff radius a to avoid this divergence"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002848_s0219581x11009350-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002848_s0219581x11009350-Figure10-1.png",
+        "caption": "Fig. 10. Schematic representation of microdisk attached to the cancer cell surface.",
+        "texts": [
+            " The saturationmagnetization,Ms, for the Permalloy is 8 105 A/m. An upper limit of the applied forces can be estimated by considering the problem in a quasi-static regime. This simpli\u00afcation can be justi\u00afed for time-varying \u00afelds with the pulse duration longer than \u00f0FV \u00de=k \u00bc 10 4 10 5 s. when no substantial energy dissipation due to viscous forces is expected. The angle between the disk plane and applied \u00afeld H is largely unknown and lies in a range of 0 90 ; therefore we compare three model cases of \u00bc 10 , 45 and 80 (Fig. 10). From the experimental and micromagnetic hysteresis loops we \u00afnd that a \u00afeld of 0.01T (100Oe) induces an in-plane component of the magnetization of 0.2, 0.1 and 0.05 of Ms for 10 , 45 , and 80 , respectively. This results in a total induced magnetic moment mper disk of 7:54 10 15, 3:77 10 15 and 1:88 10 15 Am2 for each model case. The magnetic torque m Hsin ( ) can be estimated as follows: 7:54 10 15Am2 0:01T sin\u00f010 \u00de \u00bc 1:31 10 17Nm ; 3:77 10 15Am2 0:01T sin\u00f045 \u00de \u00bc 2:67 10 17Nm ; 1:88 10 15Am2 0:01T sin\u00f080 \u00de \u00bc 1:85 10 17Nm : Therefore, the force applied to the cell at the disk edge can then be estimated as: 26, and 53 and 37 pN for \u00bc 10 , 45 and 80 , respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000209_156856106779024418-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000209_156856106779024418-Figure2-1.png",
+        "caption": "Figure 2. H-adhesion test sample (ASTM D4776).",
+        "texts": [
+            " Each 1000 m of these cords had a weight of 1 kg. First, the rubber parts were placed in the channels of a stainless steel die. Then the RFL-coated PET cords were embedded in the rubber. The procedure is shown in Fig. 1. The cords were fixed at one end and stretched using a 50 g weight. Then new rubber parts were placed on coated cords (cords were sandwiched between two rubber parts). Samples were vulcanized at 130, 140, 150 and 160\u00b0C, for a duration of 20 m each at a pressure of 35 MPa. They were then cut into H-shape samples, see Fig. 2. Static adhesion (tensile strength) was evaluated by a Monsanto 500 tester, at an elongation rate of 120 mm/min. The shear strengths (N/m2) to separate the cord from the rubber were evaluated at 25, 50, 100, 125 and 170\u00b0C. Table 1. Rubber compound composition [12, 13] Component g wt% NR (standard Malaysian rubber) 45 22 SBR 55 26.5 ZnO 5 2.4 Stearic acid 1.5 0.7 Carbon black 330 35 17 Carbon black 550 35 17 Antioxidant (4010) 3 1.5 Oil (aromatic oil 840) 15 7.3 TMTD (Tetramethyl Thiuram Disulphide) 7 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure8-1.png",
+        "caption": "FIGURE 8. The wedge (Mechanics 2.15). Drachmann\u2019s drawing (The Mechanical Technology, p. 72) is from Ms L.",
+        "texts": [
+            " If we imagine detaching the rightmost section of the weight Z from the sections G\u030cBT. , it will hold those sections in equilibrium. Thus a force equal to \u00bc the total weight of Z, applied at K, balances \u00be the total weight (G\u030cBT. ), and a slightly larger force will move it.36 In the case of the wedge and screw the similarity to the concentric circles is much less clear. Since Heron claims that the screw is simply a twisted wedge (2.17), I shall concentrate here on the analysis of the wedge in 2.15 (see Fig. 8).37 The argument is in two stages: (1) First Heron claims that any blow, even if it is slight, will move any wedge. The idea is to divide 35 Mechanics, 2.12 (Opera, vol. II, p. 126.1\u20135); Engl. trans. Drachmann, The Mechanical Technology, p. 70, modified. 36 In general, letting F represent the moving force, W the weight, and n the total number of segments of rope that bear the weight, we have F:W :: 1:n. 37 Note however that the analysis of the screw as a twisted wedge is supported by explicit reference to a practical procedure for making a screw, viz",
+            " Thus just as in the concentric circles, the effect of a small force can equal that of a large one; the difference is that in the circles the effectiveness of a force depends on its distance from the centre, while in the wedge it depends on the distance over which the force acts. Similarly, in the case of the screw Heron notes that a screw with tighter threads will be able to move a larger load by means of the same force, but it will 39 Mechanics, 2.15 (Opera, vol. II, p. 134.21\u201331); Engl. trans. Drachmann, The Mechanical Technology, p. 73, modified. 40 Alternatively, in Fig. 8, if the force BG\u030c is applied to the whole wedge MD, the load will be displaced by MD in a given time t. But if the force BH, equal to \u00bc BG\u030c, is applied to the sub-wedge RD, then the load will be moved by \u00bc MD in time t and by MD in 4t. require a greater time to do so.41 The key idea in the analyses of both the wedge and the screw is thus that of compensation between forces and the times (and distances) over which they act: if we reduce the force, we must increase the time (see further below, pp"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003836_1.3552507-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003836_1.3552507-Figure2-1.png",
+        "caption": "FIGURE 2: experimental setup",
+        "texts": [
+            " The simulation of the whole rolling process and the calculated results were developed within a DFG-funded joint project with an experimental-numerical determination of the tool stresses during cross rolling of high gearings together with the Faculty of Mechanical Engineering, Department Mechanics of Solids, Division Experimental Mechanics of the TU Chemnitz. On side of research a method for detecting tool mechanical stresses was to develop during the forming process, which were on the one hand calculated by using FEM for adaptation of tool load and on the other hand determined by experimental procedure with measurement of strain gauges. The recorded readings were transferred to the computer technology with a multi-channel wireless telemetry system from the rolling machine (see figure 2). The combination of practical determined values and simulation led to the realization of a simulation tool for process and product development, and finally to the design and optimization of cross-wedge tools with applicability to future profile geometries. The method of visioplasticity is characterized by continuum mechanic analysis of plastic deformation and flow processes. It combines the advantage of a relatively \"small\" device of technical effort with the possibility to investigate metallic materials in the quite-field"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure4.11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure4.11-1.png",
+        "caption": "Figure 4.11 \u2018s\u2019 plane representation of the first-order lag response",
+        "texts": [
+            " The negative j half of the \u2018s\u2019 plane is simply a mirror image of the positive half of the plane which is inherent in the mathematics and this mirror image represents an important feature in control systems in that when imaginary roots occur (implying oscillatory behavior) they always occur in complex conjugate pairs of the form \u00b1 j . This complex frequency interpretation of points in the \u2018s\u2019 plane provides a valuable insight into the effect that each of the roots of the control system transfer function have on the behavior of the system in the real world. Going back to our first-order lag step response problem, this can be represented by the \u2018s\u2019 plane diagram of Figure 4.11. The \u00d7\u2019s are defined as \u2018poles\u2019 meaning that for those specific values of \u2018s\u2019 the function y s defined by equation (4.4) above becomes infinite hence the term \u2018pole\u2019. As indicated in Figure 4.11 the pole at \u22121/T is referred to as the \u2018system pole\u2019 and the pole at the origin as the \u2018input pole\u2019 representing the step function. Extending this to the general case we can say that linear systems are typically of the form: F s = k ( sm+a1s m\u22121 +a2s m\u22122 + am ) sn+b1s n\u22121 +b2s n\u22122 + bn Factorizing we obtain: F s = k s\u2212z1 s\u2212z2 s\u2212zm s\u2212p1 s\u2212p2 s\u2212pn The roots of the numerator z1 z2 zm are called the \u2018zeros\u2019 of the system since when s= z the term equals zero and the function F s goes to zero. The roots of the denominator p1 p2 pn are the poles as described in the first-order lag example above so that when s = p the function F s goes to infinity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001362_tro.2008.2008745-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001362_tro.2008.2008745-Figure5-1.png",
+        "caption": "Fig. 5. Experimental results. (a) Initial configuration, (b) desired configuration, (c) initial image, and (d) desired image.",
+        "texts": [
+            " Moreover, since the joints are prismatic, it is easy to measure their offsets manually with millimetric accuracy. This is also sufficient to ensure that the gain is accurate enough. Now, to totally determine the interaction matrices, the attachment points Aj have to be computed. In [6], a calibration procedure was proposed, using leg observation. This method can be combined with the automatic leg detection to make it more practical. The proposed approach has been validated on the commercial DeltaLab Table de Stewart shown in Fig. 5. The legs of the platform have been modified to improve image processing. The experimental robot has an analog joint position controller interfaced with Linux-RTAI.Joint velocity control is emulated through this position controller with an approximate 20-ms sampling period. The omnidirectional camera used is a parabolic mirror combined with an orthographic lens. It is approximately placed at the base center. In a first experiment, a sequence of end-effector poses was performed by the robot. Further, nearly 1700 images were acquired while the robot was moving between the various poses in order to get a smooth leg\u2019s edges"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001151_pesc.2008.4591976-Figure13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001151_pesc.2008.4591976-Figure13-1.png",
+        "caption": "Figure 13. Three-dimensional view of the normalized CWT modulus of the stator phase current of the faulty machine for scales 0.03-0.05, fs=51.8Hz and n = 1470 r.p.m. (experimental results).",
+        "texts": [],
+        "surrounding_texts": [
+            "The authors thank the Research Committee of the University of Patras and particularly the Program KARATHEODORI, for funding the above work. REFERENCES [1] El Hachemi Benbouzid, M., \"A review of induction motors signature analysis as a medium for faults detection\", IEEE TRANSACTIONS ON INDUSTRIAL ELECTRONICS, Vol.47, No.5, pp.984-993, October 2000. [2] Mohamed El Hachemi Benbouzid, Senior Member IEEE and Gerald B.Kliman, Life Fellow IEEE, \"What Stator Current Processing-Based Technique to Use for Induction Motor Rotor Faults Diagnosis?\", IEEE TRANSACTIONS ON ENERGY CONVERSION, Vol.18, No.2, pp.238-244, June 2003. [3] Kral, C., Habetler, T. G., and Harley, R. G., \"Detection of mechanical imbalances of induction machines without spectral analysis of time-domain signals\", IEEE TRANSACTIONS ON INDUSTRY APPLICATIONS, Vol.40, No.4, pp.1101-1106, July 2004. [4] Mohamed Drif and Antonio J.Marques Cardoso, \"Instantaneous Real Power Signature Analysis as a Tool for Airgap Eccentricity Diagnostics in Three-Phase Induction Motors\", in Conf.Rec.ICEM 2006, Chania, Crete Island, Greece, September 2006. [5] Andreas Stavrou, \"Impedance Vector Monitoring strategy for online detection of eccentricity in induction motors\", in Conf.Rec.ICEM 2006, Chania, Crete Island, Greece, September 2006. [6] Guillermo Bossio, Student Member IEEE, Cristian De Angelo, Member IEEE, Jorge Solsona, Senior Member IEEE, Guillermo O.Garc\u00eda, Senior Member IEEE, and and Mar\u00eda I.Valla, Senior Member IEEE, \"Application of an Additional Excitation in Inverter-Fed Induction Motors for Air-Gap Eccentricity Diagnosis\", IEEE TRANSACTIONS ON ENERGY CONVERSION, Vol.21, No.4, pp.839-847, December 2006. [7] Ioannis P.Tsoumas, Epaminondas D.Mitronikas, and Athanasios N.Safacas, \"Detection of mechanical faults in asynchronous machines based on wavelet analysis of the stator current\", in Conf.Rec.ICEM 2006, Chania, Crete Island, Greece, September 2006. [8] Ioannis P.Tsoumas, \"Dynamic analysis of the subsynchronous cascade drive and development of fault diagnosis methods\" (in Greek), Dissertation, University of Patras\", August 2007. [9] Ioannis P.Tsoumas, Student Member IEEE, George Georgoulas, Epaminondas D.Mitronikas, Member IEEE, and Athanasios N.Safacas, Member IEEE, \"Asynchronous Machine Rotor Fault Diagnosis Technique Using Complex Wavelets\", accepted for publication in a future issue of IEEE TRANSACTIONS ON ENERGY CONVERSION. [10] Thomson, W. T. and Barbour, A, \"On-line current monitoring and application of a finite element method to predict the level of static airgap eccentricity in three-phase induction motors\", IEEE TRANSACTIONS ON ENERGY CONVERSION, Vol.13, No.4, pp.347-357, December 1998. [11] Thomson, W. T., Rankin, D, and Dorrell, D. G, \"On-line current monitoring to diagnose airgap eccentricity in large three-phase induction motors-industrial case histories verify the predictions\", IEEE TRANSACTIONS ON ENERGY CONVERSION, Vol.14, No.4, pp.1372-1378, December 1999. [12] Subhasis Nandi, Member IEEE, Shehab Ahmed, Student Member IEEE, and Hamid A.Toliyat, Senior Member IEEE, \"Detection of Rotor Slot and Other Eccentricity Related Harmonics in a Three Phase Induction Motor with Different Rotor Cages\", IEEE TRANSACTIONS ON ENERGY CONVERSION , Vol.16, No.3, pp.253-260, September 2001. [13] Subhasis Nandi, Member IEEE, Raj Mohan Bharadwaj, Member IEEE, and Hamid A.Toliyat, Senior Member IEEE, \"Performance Analysis of a Three-Phase Induction Motor Under Mixed Eccentricity Condition\", IEEE TRANSACTIONS ON ENERGY CONVERSION, Vol.17, No.3, pp.392-399, September 2002. [14] J.Faiz, I.Tabatabaei, and E.Sharifi-Ghazvini, \"A precise electromagnetic modeling and performance analysis of a three phase squirrel cage induction motor under mixed eccentricity condition\", Electromagnetics, Vol.24, No.6, pp.471-489, 2004 [15] P.Addison, \"The illustrated wavelet transform handbook, Institute of Physics Publishing\", 2002."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000525_s00170-006-0874-y-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000525_s00170-006-0874-y-Figure2-1.png",
+        "caption": "Fig. 2 Tooth flanks of a toothed wheel obtained with hobbing",
+        "texts": [
+            " The computer simulation of the gear generation takes place in a three dimensional mechanical desktop environment. A tooth profile is formed due to the turning of the gear wheel and the tool. The helix results from the hob blade moving along the wheel axis. The rotary motion of the tool \u03d5 2 is dependent on the rotary motion of the generated gear wheel \u03d5 1 in accordance with the transmission ratio of the technological gear, equation (1). Shift c corresponds to the value of the axial feed per wheel turn. Examples of the flanks of the teeth obtained through hobbing simulation is presented in Fig. 2. The flanks of the toothed wheel are composed of a number or regularly distributed surfaces corresponding to the successive layers of the material being removed by the tool blades. The surfaces are concave, therefore the tooth flanks are scaly in character (Fig. 3). It is evidenced by a curvature analysis in CAD of the tooth flanks (Fig. 4). Their depth along the height of the teeth varies. 3 Experimentation 3.1 Workpiece details The characteristic of the gear subjected to tests were introduced in Table 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000574_j.triboint.2008.02.010-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000574_j.triboint.2008.02.010-Figure1-1.png",
+        "caption": "Fig. 1. Schematic view of system for measuring NRRO.",
+        "texts": [
+            " ), kanada@kanto-gakui- was manufactured using the basic design concept of the NRRO measuring equipment for a spindle motor of a hard disk drive (computer peripheral parts). Then, the NRRO of an actual ball bearing was measured and the NRRO was analyzed. Furthermore, recognizing the present state of the NRRO for rolling bearings, the authors propose a procedure for improving the NRRO and demonstrate several results in this paper. A schematic view of the developed measuring system for NRRO is shown in Fig. 1. The structure is similar to that of the NRRO measuring equipment for rolling bearings used in hard disk drives [2]. This measuring system is based on the measuring methods for rolling bearings defined in ISO 1132-2:2001. The characteristics of the NRRO measuring system are as follows. (1) Axial force is loaded on a static air pressure pad. Then, the force can be loaded with non-contact. (2) If the rolling bearing to be measured is eccentric to the rotation center of the air spindle for supporting the rolling bearing, the measured data are not affected by the eccentricity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003749_s1068366612020110-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003749_s1068366612020110-Figure1-1.png",
+        "caption": "Fig. 1. Model A. Diagram of contact (a) and mechano rolling (b) fatigue tests.",
+        "texts": [
+            " The aim of the study is to develop a method for the experimental study of the influence of the stresses 1 A active system is a mechanical system in which the friction pro cess develops in all its manifestations and which simultaneously carries and transmits an alternating load (GOST 30638\u201399. Tribo Fatigue: Terms and Definitions). induced by the off contact cyclic load on the coeffi cient of resistance to rolling in a shaft\u2013roller active system, which simulates to a certain extent the basic conditions of the operation of the rail\u2013wheel system. The tests to study the regularities of the effect of cyclic stresses on friction force variation during rolling were carried out using the arrangements shown in Fig. 1 (model A) and (model B). In the mechano roll ing fatigue tests (see Figs. 1b and 2b), a cylindrical specimen\u2014a shaft fixed in the spindle of the test machine\u2014rotates with the angular velocity \u03c91. The bending load Q is applied to its free end; it is directed upwards or downwards. A counterspecimen\u2014the roller\u2014is pressed to the specimen under bending in the working zone by the contact load FN. Depending on the direction of the bending force Q, friction occurs either in the compression zone or in the tension one"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000927_2009-01-0328-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000927_2009-01-0328-Figure3-1.png",
+        "caption": "Figure 3: The equivalent curvilinear ellipsoidal solid",
+        "texts": [
+            " These are obtained by consideration of principal radii of bodies in contact at any instant of time [11]. Each contacting surface has two such principal radii; one in the direction of the normal plane and the other along the conforming flanks for a helical gear teeth pair (tangential plane yz (see figure 2). For the xz plane, the equivalent radius is: (10) Noting that: (11) and: (12) For the yz plane a partially conforming contact exists, in which the radius of the wheel is considered as negative due to its concavity, thus: (13) Therefore, as shown in figure 3, there is an equivalent ellipsoidal solid contacting/impacting a semi-infinite elastic half-space at any instant of time. The equivalent radius is that of an ellipsoidal solid impacting/contacting a semi-infinite elastic half-space of reduced elastic modulus: (14) Under rattle conditions a pair of teeth impact and rebound/separate, when the maximum Hertzian penetration is reached. For each pair of impacting teeth, represented by an ellipsoidal solid mass, m, in localised Hertzian impact (below their modal behaviour), principle of conservation of energy is upheld, when: (15) where m is the mass of the equivalent ellipsoidal solid: 2 xm V r l l is the arc length of the equivalent ellipsoidal solid, which can be approximated by the length of the contact footprint"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002341_j.optlastec.2011.04.010-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002341_j.optlastec.2011.04.010-Figure5-1.png",
+        "caption": "Fig. 5. Temperature fields in cros",
+        "texts": [
+            " In spite of these fair deviations, the calculated values are generally in good agreement with the experimental results in test cases, so that the numerical models established in present work are suitable for describing the LSM process reasonably. The substrate dimensions and LSM simulations parameters are listed in Table 4. The effects of ambient conditions on the temperature and residual thermal stress fields are discussed, respectively. s RMZ measurements in cross section for (a) case 1, (b) case 2 and (c) case 3. Sample dimensions (mm) Laser power (kW) Scanning rate (m/min) Spot radius (mm) 25 20 8 1 1.2 1 Fig. 5 gives the contours of transient temperature fields in cross section of the sample under different ambient conditions. It is seen that the effects of water assist (cases 2 and 3) are not notable and the small variations of temperature field are compared in Fig. 6. Fig. 6 shows the nodal thermal cycles for different cases in the RMZ with respective distance of 1.5 mm and 1 mm in y\u2013z cross section away from fusion line. From Fig. 6(a), it is seen that when the LSM is processed in air (case 1) and under water film (case 2), the transient temperature fields are almost the same"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003914_j.proeng.2011.05.083-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003914_j.proeng.2011.05.083-Figure5-1.png",
+        "caption": "Fig. 5. Coil arrangements of MSBS (from Sawada and Kunimasu [3])",
+        "texts": [
+            " We analyze the video images to determine the horizontal and vertical velocity components of the arrow at these two locations, from which we can estimate the drag-coefficient CD assuming that no lift force is acting on the arrow, i.e., the arrow flies ideally with zero angle of attack (see Suzuki et al. [2] for the details of analysis). The Magnetic Suspension and Balance System (MSBS) provides support-interference-free measurements, because the force supporting the arrow is generated by magnetic fields controlled by coils arranged outside the test section (Figure 5) [3]. We insert cylindrical permanent magnets inside the arrow shaft. The arrow is supported to be still against the gravitational and aerodynamic forces, by adjusting the electric currents. Then, we know inversely the aerodynamic forces acting on the arrow from the values of electric currents. The maximum speed (45m/sec) of the wind tunnel is less than the speed of the arrow launched from the crossbow (about 60m/sec). We use both the real arrow and a model which is 2.05 times larger, so that we can investigate a wider Re number range (0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002371_s00170-010-2582-x-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002371_s00170-010-2582-x-Figure3-1.png",
+        "caption": "Fig. 3 Cutting location on side region Fig. 4 Cutting location on fillet region",
+        "texts": [
+            " And the included angle 8\u2014between vectors, the tool axis and n (normal to the cutter edge at point P)\u2014is always a constant equal to \u03b1. Besides 8 and values of point Q in y- and z-directions are all equal to zero when the cutting location occurred at the tool tip (\u03b8=\u03b8Q), but this condition should be avoided due to zero cutting velocity. 2.2 Case 2: contact point P on side region Point E is the intersection of lines CD and QB. Because yE (the value of point E in y-direction) is equal to d/2, the value of point E in z-direction is solved from Eq. 4 as zE \u00bc d 2 tan a With locating along line CD as shown in Fig. 3, the cutting contact points are always restricted by tan b \u00bc yP yE zP zE \u00f06\u00de and 8 is a constant equal to p=2 b.Fig. 2 Cutting location on bottom region Fig. 1 Geometric parameters about milling tool Point P in y- and z-directions are also solved by substituting Eqs. 6 into 3 and can be expressed in terms of h as zP \u00bc fh2 2e d 1 tan a tan b\u00f0 \u00de\u00bd h f h2 2 tan bh 1 yP \u00bc f tan b \u00fe d 1 tan a tan b\u00f0 \u00de=2\u00bd h2 2e tan bh f tan b \u00fe d 1 tan a tan b\u00f0 \u00de=2\u00bd h2 2 tan bh 1 Since the value of point D in z-direction is L, yD can be obtained from Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-18-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-18-1.png",
+        "caption": "Figure 3-18: Wear due to fretting corrosion.",
+        "texts": [
+            " The heavy specific load imposed on that part of the ring immediately over the turning chip produced the premature spalling seen in the right-hand illustration. On either side of the spalled area is a condition called fragment denting, which occurs when fragments from the flaked surface are trapped between the rollers and the raceway. When the contact between a bearing and its seat is not intimate, relative movement results. Small movements between the bearing and its seat produce a condition called fretting corrosion. Figure 3-18 illustrates a bearing outer ring that has been subjected to fretting corrosion. Figure 3-19 illustrates an advanced state of this condition. Fretting started the crack, which in turn triggered the spalling. Figure 3-20: Fretting caused by yield in the shaft journal. Machinery Component Failure Analysis 107 Fretting corrosion can also be found in applications where machining of the seats is accurate, but where on account of service conditions the seats deform under load. Railroad journal boxes are an example of this condition"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000809_robot.2008.4543556-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000809_robot.2008.4543556-Figure6-1.png",
+        "caption": "Fig. 6. Overview of the experimental setup",
+        "texts": [
+            "3, where the most efficient value of a wheel traction is obtained comparing to our previous researches. In order to verify the proposed control, slope traversal experiments were carried out using a rover test bed developed in our research group. In particular, to generate larger dynamic slip motions of the rover, we conducted the experiments on the side slope traversal situation using a tiltable test field. The performance of the proposed control is evaluated based on the distance and orientation errors. Fig. 6 shows the overview of the experimental setup with our rover test bed on the tiltable test field. The test field consists of a flat rectangular soil-vessel in the size of 2.0 by 1.0 [m]. The vessel is filled up with 8.0 [cm] depth of a Toyoura Sand, which is loose sand and standard sand for terramechanics research filed. The vessel can be inclined up to 20 [deg]. 1) Rover test bed: The four-wheeled rover test bed as shown in Fig. 7 has a dimension of 0.44 [m] (wheelbase) \u00d7 0.30 [m] (tread) \u00d7 0.30 [m] (height) and weights about 13"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003776_j.mechmachtheory.2012.03.003-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003776_j.mechmachtheory.2012.03.003-Figure2-1.png",
+        "caption": "Fig. 2. Identical symmetric general tori formed by various circles with parameters r1, s1(r1> s1),\u03c71 (A), with parameters r2, s2(r2b s2),\u03c72 (B).",
+        "texts": [
+            " The Bennett linkage will be formed if the axes of the circles that pass through points Q, R1, R2, and S are revolute pairs which connect links QR1=r1, R1S=s1, QR2=r2, R2S=s2. In a kinematic system based on the radii of circles forming a torus to be a Bennett linkage, conditions need to be fulfilledwhen the revolute pair at point S allows for complete rotational movements. These conditions can be defined as follows: \u03c71=\u03b62=\u03c0/2,\u03b61=\u03c72. There are two symmetric mechanisms with \u03c71=\u03c0/2 depending on the sign (+ or \u2212) of angle \u03c72. Circle with the radius si and axis RiZi are located on the plane XYi that is turned around an axis QX by angle \u03c71 (Fig. 2). Rotation of such a circle around an axis QZ, where ri is the distance between the center of the circle and an axis, forms a surface of general symmetric torus. If the angle \u03c71, which defines the position of the plane XYi in relation to axis QY, is equal to \u03c0/2, the torus becomes a classic torus. If the plane is additionally rotated around an axis QYi, the general (asymmetric) torus is obtained [16,20]. Similarly, when Villarceau circles are located on the classic torus, there are two different circles. A symmetric general torus is formed by the rotation of these circles, (Fig. 2). An analysis of such a case is presented below. Two identical tori can be obtained for various sets of parameters. In Fig. 2(A, B), the points located on torus surface Si and unit vectors Ui, which are located on planes perpendicular to radius si of the circle: Si \u00bc 1 0 0 0 c\u03c7i \u2212s\u03c7i 0 s\u03c7i c\u03c7i 2 4 3 5 c\u03d5i \u2212s\u03d5i 0 s\u03d5i c\u03d5i 0 0 0 1 2 4 3 5 si 0 0 2 4 3 5; \u00f01\u00de Ui \u00bc 1 0 0 0 c\u03c7i \u2212s\u03c7i 0 s\u03c7i c\u03c7i 2 4 3 5 c\u03d5i \u2212s\u03d5i 0 s\u03d5i c\u03d5i 0 0 0 1 2 4 3 5 1 0 0 0 c\u03b6 i \u2212s\u03b6 i 0 s\u03b6 i c\u03b6 i 2 4 3 5 0 0 1 2 4 3 5: \u00f02\u00de Once the vectors have been shifted by radius ri and turned around axis QZ by angle \u03c3i, we obtain equations in an immovable coordinate system [17]: S\u03c3i \u00bc S\u03c3ix S\u03c3iy S\u03c3iz 2 4 3 5 \u00bc sic\u03d5i \u00fe ri\u00f0 \u00dec\u03c3 i\u2212sis\u03d5c\u03c7is\u03c3 i sic\u03d5i \u00fe ri\u00f0 \u00des\u03c3 i \u00fe sis\u03d5ic\u03c7ic\u03c3 i sis\u03d5is\u03c7i 2 4 3 5 \u00f03\u00de U\u03c3 i \u00bc U\u03c3 ix U\u03c3 iy U\u03c3 iz 2 4 3 5 \u00bc s\u03b6 is\u03d5ic\u03c3 i \u00fe s\u03b6 ic\u03d5ic\u03c7i \u00fe c\u03b6 is\u03c7i\u00f0 \u00des\u03c3 i s\u03b6 is\u03d5is\u03c3 i\u2212 s\u03b6 ic\u03d5ic\u03c7i \u00fe c\u03b6 is\u03c7i\u00f0 \u00dec\u03c3 i \u2212s\u03b6 ic\u03d5is\u03c7i \u00fe c\u03b6 ic\u03c7i 2 4 3 5 \u00f04\u00de Vector (3) describes the location of points Sion the surface of a symmetric torus, and Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003971_icra.2012.6225012-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003971_icra.2012.6225012-Figure4-1.png",
+        "caption": "Fig. 4: This Figure shows the importance of deflections during grasping. To the left, we show the gripper position and configuration of bar if there are no deflections. In the middle, we show the same position of the gripper but now with a visualization of the deflected bar. To the right, we show the adjusted position of gripper when the deflection is accounted for in the offline programming.",
+        "texts": [
+            " The main steps in the robotic system are: \u2022 bend the bar into the 3D shape described by the task input \u2022 calculate the grasping point on the deflected bar and grasp the bar at that point \u2022 transport the bar to the reinforcement structure and use the deflection to place the bar correctly on the structure \u2022 bind the bar to the structure A reinforcement bar can be bended to an arbitrary 3- dimensional shape at the bending station as described in the previous section. Next the bended bar needs to be picked up by the robot. This process is delicate as hinted in the previous sections since the deflection of the bar may be significant at the desired grasping point. This is illustrated in Figure 4 where the middle image show how the grasping would fail if the calculation of the grasping point does not take the deflection into account. The right most image illustrate the successful grasp point when deflection is included in the calculation. It is important that the expected position and orientation of the bar at the grasping position is sufficiently accurate so that only small corrections (1-2cm or so) to the offline programmed path are necessary at runtime. These corrections can be compared to e"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000624_00405000701592975-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000624_00405000701592975-Figure1-1.png",
+        "caption": "Figure 1. Projection of a fibre between two bonds on XY plane.",
+        "texts": [
+            " Since, the average bond\u2013bond distance is considered to be small such that the fibre segment be- tween the two successive bonds can be assumed as a straight line. Torsional and tensile deformations of the fibre segments are so small that they can be neglected. Fibres do not slip past each other, i.e. frictional characteristics between the fibres have been neglected. Consider a fibre of average length b\u0304 between the two bonds whose direction with respect to spherical co-ordinate system is defined by out of plane (\u03b8 ) and in-plane (\u03d5) fibre orientation angles, as shown in Figure 1. Assuming, the fibre orientation distribution function, , describing the probability of finding a fibre in the infinitesimal range of angle \u03b8 and \u03b8 + d\u03b8 and \u03d5 and \u03d5 + d\u03d5, is defined by (\u03b8, \u03d5) sin \u03b8 d\u03b8 d\u03d5. However, the following normalisation condition must be satisfied: \u03c0\u222b \u03b8=0 \u03c0/2\u222b \u03d5=\u2212\u03c0/2 (\u03b8, \u03d5) sin \u03b8d\u03b8 d\u03d5 = 1. (1) According to Komori and Makashima (1977) and Lee and Lee (1985), the mean length of the fibre between the centre of two neighbouring bonding points projected on the planar direction, b\u0304j (here the subscript j represents a planar direction), can be expressed as follows: b\u0304j = b\u0304Kj = V 2DLI Kj = \u03c0D 8Vf I Kj (2) where D is the fibre diameter and L is the total length of the fibres in a volume V , Vf is the fibre volume fraction, I is an integral defining the orientation distribution function of D ow nl oa de d by [ A st on U ni ve rs ity ] at 0 3: 20 2 1 Ja nu ar y 20 14 the fibres and Kj is the directional parameter or geometric coefficient when b\u0304 is projected on the planar direction, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001686_09596518jsce623-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001686_09596518jsce623-Figure9-1.png",
+        "caption": "Fig. 9 Schematics of the section of a couple of RGBs: top view",
+        "texts": [
+            " 8(a), are installed between the drives and the proximal P joints. Their main function is the generation of the tilt motion from the rotation of the planets 6 with respect to their planet carriers 8. The gear ratio of the RGB is 1:1, and hence this subsubsystem does not have any effect on the joint kinematics of the system. However, its inertia has to be taken into account as its parts are made of steel. This subsubsection aims at deriving the contribution of the RGB to the inertia and the C-matrices of the system. Figure 9 depicts a layout of one pair of RGBs, the one driving the distal P joint of limb I. Each pair, enclosed in a housing H, is composed of (a) two vertical bevel gears, turning at the rates v6\u2013v8 with respect to their planet carrier 8, as these are rigidly coupled to two of the three planets, and (b) its two meshing horizontal bevel gears, turning at the same rates. The inertia of each vertical bevel gear has already been included in that of its planet gear 6, the one of the housing having been included in that of the planet carrier 8, to which it is rigidly attached"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-204-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-204-1.png",
+        "caption": "Figure 3-204: Typical arrangement of Sohre Turbomachinery, Inc. \u201ctoothbrush\u201d type brushes running on shaft OD. These brushes can also be run against a shaft end of collar when equipped with spring assist.",
+        "texts": [],
+        "surrounding_texts": [
+            "using a reliable gaussmeter with a hall probe, then the need for demagnetizing is clear. Following this, efforts should be made to maintain the equipment in a demagnetized state. Some of the suggestions that may be put into action are discussed below. 1. Existing Installations For the many units now installed and being manufactured, the prospect of residual magnetism causing problems is very real. Units should be maintained in a demagnetized state. Every effort should then be made to control and prevent remagnetization. \u2022 All components upon receipt following purchase, repair, or testing should be entirely free of residual magnetism. \u2022 Thorough deep-soaked demagnetization should be conducted on any component following magnetic particle inspection or on any component discovered to have high level of magnetic fields. Machinery Component Failure Analysis 291 \u2022 Welding on the compressor, turbine, or its piping should be controlled very carefully. The ground clamp and electrode cables should both be strung along the same path to the work area. Then the ground clamp should be connected to the same metal piece that is to be welded. \u2022 All components should have ground straps interconnected to the structure or station ground grid. The ground grid should have a ground resistance of less than 3 ohms. Also, lightning rods and other tall structures should have cables firmly interconnected to the ground grid. They further must be routed so they are not near nor do they link magnetic circuits such as closed-loop piping between the compressors or turbine to heat exchangers, condensors, boilers, etc. The goal is to provide a low-impedance discharge path for atmospheric discharge current but in such a way that component magnetization cannot occur. \u2022 Reliable brushes should be applied to the shafts to drain away electrostatic charge and to shunt persistent electromagnetic currents around bearing and seal surfaces. Brushes must be continuously conducting. This may require the use of a properly designed wire bristle brush (Figure 3-205). Close initial monitoring of the brush currents and voltages is necessary to assure that there is no compounding of magnetic fields due to the internal current paths. If this occurs, then brushes should be removed until components are demagnetized and there is assurance that remagnetizing currents are arrested. 2. Future Installations Remagnetization and occurrence of bearing damage is expected to continue unless significant measures are taken to correct the problems. Some of these measures are: \u2022 Insulation of all bearings, seals, couplings, and other components through which damaging currents now flow. \u2022 Installation of permanent brushes to shunt currents around the affected components. \u2022 Selection of materials for the equipment that is magnetically soft rather than magnetically hard. \u2022 Installation of coils in units at the time of original manufacture that can be used to effect demagnetization without having to disassemble the equipment."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002596_iros.2010.5651607-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002596_iros.2010.5651607-Figure3-1.png",
+        "caption": "Fig. 3. Simulation model",
+        "texts": [
+            " On the other hand, identification of faulty channel sometimes failed when the channel number exceeds measurement DOF only 1. Although this study uses a multi support mechanism with 3 force sensors, it is also applicable to a force sensor unit with redundant channels of strain measurement. This section presents the results obtained from simulations conducted to verify the proposed algorithms. Another purpose of the simulation is to verify the error from the true value, which cannot be measured in experiments. The simulation models used for the purposes are shown in Fig. 3. Here, the simulations are performed on a two-dimensional plane for simplification. In the model structure, a circular end-effector is supported by three 2-axis force sensors. The three force sensors are fixed to the center base and are used for measurement of forces in the x\u2212/y\u2212axis directions. The end-effector moves slightly in the x\u2212/y\u2212axis directions as well as in its rotating direction and therefore can be utilized as a 3-axis force/torque sensor. Its angle of rotation is \u03b8. A total of six channels (2-axis\u00d73) can be measured by this end-effector, and it has a redundancy corresponding to three degrees of freedom"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure5.10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure5.10-1.png",
+        "caption": "Fig. 5.10-1 A double-rod pendulum [5.16].",
+        "texts": [
+            "indb 166 11/24/10 11:46:32 AM Multi-Degree-of-Freedom Systems | Chapter 5 167 At this point, it is convenient to introduce a Lagrangian function L as L = \u2212T U (5.10-12) By observing that U is not a function of the qr, we can write Eq. 5.10-11 as d dt q qr r \u2202 \u2202 \u239b \u239d\u239c \u239e \u23a0\u239f \u2212 \u2202 \u2202 =L L 0 (r n= \u22c5\u22c5\u22c51 2, , , ) (5.10-13) Eq. 5.10-13 represents the standard form of Lagrange\u2019s equation for conservative systems. Example 5.10-1 [5.16] Two identical pin-connected and pin-supported uniform rods are moving in a vertical plane, as shown in Fig. 5.10-1. Let the orientations of the rods be defi ned by the angles 1 and 2, as shown. Let each rod have length b and mass m. Assume the pin connections are frictionless. Use the Lagrangian method to derive the diff erential equations of motion and obtain the natural frequencies for small vibration by assuming small oscillation. Solution: Let the rods be called B1 and B2, and let their mass centers be G1 and G2. Introducing unit vectors n11, n12, n21, and n22, as shown, the mass center velocities of the rods can be expressed as V nG b1 2 1 11=( / ) and V n nG b b2 1 11 2 212= + ( / ) (a) Th e angular velocities of the rods are B1 1 3= n and B2 2 3= n , (b) where the unit vector n3 is normal to the plane of motion",
+            " (4) Kane\u2019s Equations Kane\u2019s equations state that the sum of the generalized inertia force and the generalized applied force of the system is zero for each generalized coordinate qr, namely, F Fq qr r * + = 0 (r n=1 2, , , ) (5.11-18) Example 5.11-1 Repeat Example 5.10-1, using Kane\u2019s method to derive the equations of motion of the system. Solution: As before, let the rods be called B1 and B2, and let their mass centers be G1 and G2. Introducing unit vectors n11, n12, n21, and n22, as shown in Fig. 5.10-1, the mass center velocities of the rods may be expressed as V nG b1 2 1 11=( / ) and V n nG b b2 1 11 2 212= + ( / ) (a) Th e angular velocities of the rods are B1 1 3= n and B2 2 3= n , (b) where n3 is a unit vector perpendicular to the plane of motion. Th e mass center accelerations and the angular accelerations of the rods are a n nG b b1 2 21 11 1 2 12= +( / ) ( / ) (c) a n n nG b b b b2 1 11 1 2 12 2 21 22 2= + + + ( / ) ( / ) 2 22n (d) B1 1 3= n (e) B2 2 3= n (f) From Eqs. a and b, the partial velocities of G1 and G2 for 1 and 2 are seen to be V n 1 1 2 11 G b= ( / ) and V 2 1 0G = (g) V n 1 2 11 G b= and V n 2 2 2 21 G b= ( / ) (h) VibrationAnalysis_txt"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure3-1.png",
+        "caption": "Figure 3. Static balancing with the rectilinear trajectories of the counterweights positioned arbitrarily.",
+        "texts": [
+            " Consequently, for the static balancing of any plane mechanism with two counterweights, it is possible to select both trajectories arbitrarily. By way of example, consider the method balancing a plane mechanism with two counterweights, moving along two mutt, ally perpendicular lines (Fig. 2). The trajectories of the centers of the counterweights correspond to the relationship x' t = a and _z=-' b. However, it is possible to select as trajectories any straight lines (with the exception of parallel lines) intersecting at any angle. Such an arrangement is shown in Fig. 3, where the trajectories are given by the relationships and z i =(x i - a)tg .x - ' = b \";2 The magnitudes a, b, and 5( must be arbitrarily selected. However, from a design point of view, it is more expedient to have a counterweight whose center of mass moves in a circle and not in a straight line. Figure 4 shows the trajectory of the mass center of the moving links and two circular trajectories of the counterweights. Let us write the equations for these trajectories: (St - x l ) 2 + (z't) 2 =p~; ( x "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003005_s0263574711001172-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003005_s0263574711001172-Figure12-1.png",
+        "caption": "Fig. 12. The 3-RRR planar parallel manipulator.",
+        "texts": [
+            "8252 times the time required by the developed approach as a minimum, since no iteration is required for the developed approach. The absolute tracking errors in both X- and Y-directions are shown in Fig. 11 in which the maximum tracking error is 0.1 mm, which is higher than the one obtained by the developed approach after applying the fine tuning stage. This manipulator is symmetric with three identical limbs connecting the equilateral triangle of the base and that of the moving end-effector as shown in Fig. 12. The vector lengths of af and bf , f = 1, . . . , 3, are 120 mm and 80 mm, respectively. The length of each side of the moving equilateral triangle is 100 mm, while that of the base is 300 mm. Point G is taken as the median point of the moving end-effector. The required end-effector pose vector is defined as x = [xG yG \u2205 ]T , while the given actuated joint angle variables vector is defined as g = [\u03c81 \u03c82 \u03c83 ]T and that of the passive joints is given by \u03b8 = [ \u03b81 \u03b82 \u03b83 ]T . The forward kinematics of this manipulator has six possible solutions while the inverse kinematics has eight feasible solutions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003324_978-3-642-33503-7_17-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003324_978-3-642-33503-7_17-Figure1-1.png",
+        "caption": "Fig. 1. Schematic of the proposed technique operation modes. (a) Mode 1 and (b) Mode 2",
+        "texts": [
+            " The previously mentioned mapping techniques are either having drawbacks like interruption of the operation, offset, reducing control accuracy or of complicated nature. So the main objective of this research is to introduce a simple and efficient technique to accomplish the task of workspace mapping of kinematically dissimilar robots in a continuous yet easy way. This method is inspired by a technique named the bubble technique used in virtual reality for spanning large virtual workspace [4]. The concept of the technique is illustrated in Fig. 1. The position of the slave robot end effector is determined according to eq. (1): (1) where Xe, Xw, Xh represent the position of the robot end effector, the virtual workspace position relative to the world frame, and the haptic device cursor position relative to the haptic device frame respectively. The constant K is a scaling factor that should be equal to or less than unity. In our case, it is chosen to be unity. The operation can be divided into two modes as in Fig. 1: \u2500 Mode 1 fine manipulation mode: in which the virtual workspace position Xw remains fixed. As moving the haptic device, the same displacement is mapped to the robot arm relative to Xw. With unity scale factor, the operator can fine manipulate a workspace volume equal to that of the master haptic device. This is typically a position control scheme. In this mode the stylus switch is released. \u2500 Mode 2 coarse motion mode: that should be switched to when it is desired to work in an area in the slave robot workspace that cannot be reached with the current value of the virtual workspace position Xw"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000183_j.jcis.2005.11.010-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000183_j.jcis.2005.11.010-Figure3-1.png",
+        "caption": "Fig. 3. Cyclic voltammograms recorded at \u03bd = 4 mV/s for 20 mM \u03b2-mAC at a GCE in different solutions: (a) 0.1 M NaBr; (b) 0.1 M NaBr + 100 mM TDTAB; (c) solution as in (b) + 2 mM [Co(II)(bpy)3]2+. Dotted line: CV of 2 mM [Co(II)(bpy)3]2+ in 0.1 M NaBr + 100 mM TDTAB.",
+        "texts": [
+            " A similar CV response is observed for [Co(II)(bpy)3]2+ in the presence of CTAB (curve (c)) and CBDAC (curve (d)); however, in the case of CBDAC there is a systematic shift in Epc and Epa to more negative potentials with E1/2 = \u22121190 mV. The negative shift in E1/2 of [Co(bpy)3]2+/+ in the presence of CBDAC surfactant is an indication that CBDAC micelles additionally stabilize the oxidized form [33], i.e., the Co(II) complex, and this may be related to additional interaction arising due to the benzyl group on \u2013N of CBDAC, a feature that is absent in CTAB and TDTAB molecules. Fig. 3 illustrates CVs recorded at \u03bd = 4 mV/s for 20 mM \u03b2-mAC at an activated GCE in 0.1 M NaBr solution under different experimental conditions. \u03b2-mAC does not undergo direct reduction on an activated GCE either in the absence (curve (a)) or in the presence of 100 mM TDTAB (curve (b)). Similarly, it is not reduced when [Co(II)(bpy)3]2+ alone is present in solution (not shown). Only when the solution contains TDTAB is \u03b2-mAC electrocatalytically reduced in the presence of [Co(II)(bpy)3]2+, curve (c). There is a considerable increase in the reduction current of the micelle-bound Co(II) to Co(I) complex at Epc = \u22121150 mV with complete absence of anodic peak, signifying mediated electrocatalytic reduction of \u03b2-mAC by the electrogenerated micelle-bound [Co(I)(bpy)3]+ regenerating the parent species, [Co(II)(bpy)3]2+"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003763_s10068-012-0158-2-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003763_s10068-012-0158-2-Figure5-1.png",
+        "caption": "Fig. 5. Response surface and contour plots showing the effect of reaction temperature and molar ratio of lauric acid to erythorbic acid on the molar conversion yield at an enzyme content of 3,000 PLU.",
+        "texts": [
+            " At a fixed enzyme content, the molar conversion yield increased rapidly when the temperature reached approximately 50oC, and it then leveled off. At a fixed temperature, the molar conversion yield was varied slightly with increasing enzyme content, especially when the temperature exceeded 50oC. This result indicates that the reaction temperature had the greatest effect on the molar conversion yield. The effect of the substrate ratio on the molar conversion yield at varying temperatures and at an enzyme content of 3,000 PLU is shown in Fig. 5. The increase in both the reaction temperature and the molar ratio of lauric acid to erythorbic acid elevated the molar conversion yield. However, increasing the reaction temperature and molar ratio of lauric acid to erythorbic acid above their optimum values (approximately 50oC and 20, respectively) did not further increase the molar conversion yield. Optimization of the synthesis of erythorbyl laurate and verification of model The conditions that resulted in the highest molar conversion yield were as follows: an enzyme content of 2,994 PLU, a molar ratio of lauric acid to erythorbic acid of 24"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003905_s00542-013-2023-5-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003905_s00542-013-2023-5-Figure1-1.png",
+        "caption": "Fig. 1 World\u2019s smallest FDB by nTn corporation (http://www.ntn.co.jp/english/ news/news_files/new_products/ news201000012.html)",
+        "texts": [
+            " among these studies, herringbone-grooved journal bearings (hGJBs) have been commonly employed because they allow lubricants to be pumped into the bearing and reduce side leakage of the oil film. none of the aforementioned works, however, put effort towards the case of the FDB designed for miniature spindle motor. however, to the best of the current authors\u2019 knowledge, for the design parameters of FDBs, their inner diameters are bigger than 1.2 mm, except for developed and commercialized the world\u2019s smallest FDB by nTn corporation for a shaft diameter of 0.6 mm by combining electrocasting and plastic injection molding technologies (see Fig. 1) (2010). To improve the herringbone-liked groove profile, there may be many parameters of groove\u2019s appearances-groove width, groove depth, groove angle, groove curvature, and groove-stepped profile. To find the optimal design of groove appearance, executing the full-factorial experimental design tests costs time. Therefore, instead of a full-factorial experimental study, a numerical study for main factors is an effective way to analyze the groove parameters. To effectively design the groove\u2019s appearances, Taguchi (1986, 1987) introduced robust parameter design that models the controllable factors in a process along with the uncontrollable or noise factors"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002169_iros.2009.5354046-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002169_iros.2009.5354046-Figure1-1.png",
+        "caption": "Fig. 1. Continuum model consisting of smooth backbone curve with zero thickness moving forward with acceleration \u03b1",
+        "texts": [],
+        "surrounding_texts": [
+            "The continuum model under the assumption that there is no lateral slippage and its approximation for rigid link model is briefly reviewed in this section. The continuum model consists of a smooth curve of zero thickness parameterized by its arc length s \u2208 [0, L]. On each point O(s) of the curve, a frame set consisting of three orthonormal bases e1, e2, e3 is attached, where e1 corresponds to the tangent of the curve O\u2032. Hereafter, derivative with respect to arc length s is denoted by by the prime sign (\u2032), time derivative by over dot (\u0307), and inner product of two vectors by \u3008\u00b7, \u00b7\u3009. Here we assume the following conditions. 1) The backbone is not stretchable (\u2016e1\u2016 = const.) 2) The backbone is torsion free (\u3008e\u20322, e2\u3009 = \u3008e\u20323, e2\u3009 = 0) 3) Lateral velocity is zero \u3008O\u0307, e1\u3009 = \u2212v, \u3008O\u0307, e2\u3009 = \u3008O\u0307, e3\u3009 = 0 where v is the progress speed along \u2212e1 4) There is no longitudinal friction 5) Rotation around the curve is not restricted except for the head 6) Bending moment \u03c42, \u03c43 respectively along e2, e3 de- fined as \u03bai = \u3008O\u2032, ei\u3009 (i = 2, 3) can be generated at arbitrary point of the body The motion of equation of the entire system can be given by m\u03b1 = \u2212 \u222b L 0 (\u03ba\u2032 2\u03c43 \u2212 \u03ba\u2032 3\u03c42) ds (1) where \u03b1 is the forward longitudinal acceleration and \u03ba2, \u03ba3 are the curvatures around e2, e3, respectively. The optimal bending moment that minimizes the quadratic cost function J = \u222b L 0 ( \u03c42 2 + \u03c42 3 ) ds can be derived by solving isoperimetric problem [6] as follows: [ \u03c42 \u03c43 ] = \u2212 m\u03b1 \u222b L 0 (\u03ba\u20322 2 + \u03ba\u20322 3 )ds [ \u2212\u03ba\u2032 3 \u03ba\u2032 2 ] (2) This continuum model can be approximated to a rigid link model by the following assignments: Curvature (\u03ba2(s), \u03ba3(s)) \u2192 ith joint angle (\u03c62[i], \u03c63[i]); bending moment (\u03c42(s), \u03c43(s)) \u2192 joint torque (\u03c42[i], \u03c43[i]) on ith joint. Thus we have [ \u03c42[i] \u03c43[i] ] = K(vd \u2212 v) [ \u03c62[i \u2212 1] \u2212 \u03c62[i] \u03c63[i \u2212 1] \u2212 \u03c63[i] ] (3) as the discrete approximation, where K is the control gain that governs the longitudinal acceleration and vd is the desired longitudinal velocity. (3) simply forms P control of the posterior joints (\u03c62[i], \u03c63[i]) whose references are the anterior joint angles (\u03c62[i \u2212 1], \u03c63[i \u2212 1]). In the two dimensional case, joint angle and torque vectors are simply substituted by scalar values \u03c6[i] and \u03c4 [i], respectively, as follows. \u03c4 [i] = K(vd \u2212 v)(\u03c6[i \u2212 1] \u2212 \u03c6[i]) (4)"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.10-1.png",
+        "caption": "Fig. 1.10. The steer angle versus lateral acceleration at constant path curvature (left graph). The difference in slip angle versus lateral acceleration and the required steer angle at a given path curvature (right graph). The understeer gradient q.",
+        "texts": [
+            "53) m g aC1- be2 s m g C (1.54) l C1C 2 1 C1C 2 with g denoting the acceleration due to gravity. After having defined the lateral acceleration which in the present linear analysis equals the centripetal acceleration: V 2 a - V r - (1.55) Y R Eq.(1.53) can be written in the more convenient form / ay J - / l + r / - - - + r / ~ (156) R R g The meaning of understeer versus oversteer becomes clear when the steer angle is plotted against the centripetal acceleration while the radius R is kept constant. In Fig. 1.10 (left-hand diagram) this is done for three types of vehicles showing understeer, neutral steer and oversteer. Apparently, for an understeered vehicle, the steer angle needs to be increased when the vehicle is going to run at a higher speed. At neutral steer the steer angle can be kept constant while at oversteer a reduction in steer angle is needed when the speed of travel is increased and at the same time a constant turning radius is maintained. 26 TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY According to Eq",
+            " Consequently, as one might expect when the centrifugal force is considered as the external force, a vehicle acts overste~reA when the neutral steer point lies in front of the centre of gravity and underste~reA when S lies behind the c.g.. As we will see later on, the actual non-linear vehicle may change its steering character when the lateral acceleration increases. It appears then that the difference in slip angle is no longer directly related to the underste~r gradient. Consideration of Eq.(1.56) reveals that in the left-hand graph of Fig. 1.10 the difference in slip angle can be measured along the ordinate starting from the value 1/R. It is of interest to convert the diagram into the graph shown on the fight-hand side of Fig. 1.10 with ordinate equal to the difference in slip angle. In that way, the diagram becomes more flexible because the value of the curvature 1/R may be selected afterwards. The horizontal dotted line is then shifted vertically according to the value of the relative curvature 1/R considered. The distance to the handling line represents the magnitude of the steer angle. Influence o f the Pneumatic Trail The direct influence of the pneumatic trails ti may not be negligible. In reality, the tyre side forces act a small distance behind the contact centres",
+            " It can easily be observed from this diagram that relation (1.81) holds approximately when the angles are small. The ratio of the side force and vertical load as shown in (1.80) plotted as a function of the slip angle may be termed as the normalised tyre or axle characteristic. These characteristics subtracted horizontally from each other produce the 'handling curve'. Considering the equalities (1.80) the ordinate may be replaced by ay]g. The resulting diagram with abscissa al-a2 is the non-linear version of the right-hand diagram of Fig. 1.10 (rotated 90 o anti-clockwise). The diagram may be completed by attaching the graph that shows for a series of speeds V the relationship between lateral acceleration (in g units) ay[g and the relative path curvature 1/R according to Eq.(1.55). Figure 1.17 shows the normalised axle characteristics and the completed handling diagram. The handling curve consists of a main branch and two side lobes. The different portions of the curves have been coded to indicate the corresponding parts of the original normalised axle characteristics they originate from"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000046_s00453-006-1206-1-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000046_s00453-006-1206-1-Figure7-1.png",
+        "caption": "Fig. 7. The awakening tree is partitioned into heavy paths, each of which is partitioned into subpaths of length \u03be .",
+        "texts": [
+            " Thus, in any root-to-leaf path in T there are at most log n light edges. Also, heavy edges form a collection of disjoint paths (because there is one heavy edge from a node to one of its children). We say that a heavy path \u03c0 \u2032 is a child of heavy path \u03c0 if one end node of \u03c0 \u2032 is the child of a node in \u03c0 . The heavy-path decomposition forms a balanced tree of heavy paths, because any root-to-leaf walk in T visits at most log n light edges, and therefore at most log n heavy paths. We use these heavy paths to refine the description of the wake-up tree. See Figure 7. We can assume that in T each heavy path is awakened by one robot, the robot that awakens the head of the heavy path (node closest to v0) and that no robot awakens more than one heavy path. In this way, a heavy-path decomposition of T corresponds to an awakening schedule with one robot per path. Because T has makespan t , each heavy path has length at most t . We divide the heavy path into subpaths of length \u03be = \u00b5t/(2 log n). Note that on any root-to-leaf path in T , we visit at most O((1+ 1/\u00b5) log n) different subpaths"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002157_icctd.2009.202-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002157_icctd.2009.202-Figure7-1.png",
+        "caption": "Figure 7. The five points and selected point for virtual target",
+        "texts": [
+            " Virtual Target In this method, with using a virtual target in around suddenly obstacle, we try to pass from suddenly obstacle and escape from local minimum. At first, for determining the virtual target's distance, we need to r radius. The Fig 6 shows r radius. The size of r depends on obstacle's magnitude. With using this radius, we gain one point around the robot. We select this point over the sudden obstacle. Nearest point to basic target with distance r as virtual target's point, that it hasn't obstacle's cell. Fig 7 shows point and virtual target's point. Than without regarding first wave, we expand a new wave from virtual target with using first wave algorithm. this wave continues until reaching robot point and in case of reaching robot point, it stops and the robot moves in base of this new wave to virtual target ,it passes from suddenly obstacle and escapes from local minimum, then it moves according to first wave to basic target. 4. Wave expansion from obstacle At first, the sudden obstacle must be identified"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003932_1077546312461026-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003932_1077546312461026-Figure2-1.png",
+        "caption": "Figure 2. Schematic of the rudder.",
+        "texts": [
+            " Furthermore, linear viscous damping terms along with sea current and wind resistive forces and moments are computed as in Isherwood (1973), Fossen (1994), Oil Companies International Marine Forum (1994), Journe\u0301e and Massie (2001), Perez (2005) and Ueng et al. (2008). The unconstrained angular displacement of the rudder is governed by Jrud \u20ac \u00bc srud erud\u00f0 \u00de cos \u00f0 \u00deFrudy sin \u00f0 \u00deFrudx Trud cont \u00f01\u00de where Frudy \u00bc Fdrag sin e\u00f0 \u00de \u00fe Flift cos e\u00f0 \u00de, Frudx \u00bc Fdrag cos e\u00f0 \u00de Flift sin e\u00f0 \u00de and Trud cont is the rudder control torque. The geometric parameters srud, erud and 1 are defined in Figure 2. In the current derivation, the velocity vector of the fluid approaching the rudder is considered to be affected by the propeller (Lewis, 1988; Perez, 2005) and the sway motion of the ship (Khaled and Chalhoub, 2011). This term is required for evaluating the coefficients of the lift and drag forces of the rudder (Abbott and von Doenhoff, 1958; Journe\u0301e and Pinkster, 2002). Equation (1) is applicable whenever min max. Beyond this range, will be set equal to either min or max and the rudder torque will be determined from equation (1) by setting \u20ac \u00bc 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002022_1077546310384002-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002022_1077546310384002-Figure3-1.png",
+        "caption": "Figure 3. Integration of forces acting on the ith pad and on the shaft.",
+        "texts": [
+            "comDownloaded from the linear velocity of the rotor surface, Pi is the injection pressure in the ith pad (active lubrication), which can be obtained by solving the Equation (9), Fi\u00f0x, y\u00de is a positioning function of the orifices on the pad surface, and Fi x, y\u00f0 \u00de \u00bc d 2 0 =4 2, for 2 d 2 0 =4 0, for 2 4 d 2 0 =4 \u00f014\u00de where \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0x xi0\u00de 2 \u00fe \u00f0 y yi0\u00de 2 q , \u00f0xi0, yi0\u00de is the position of the orifice center on the ith pad surface. The equation of motion of the ith pad can be written as Is \u20ac i \u00bcM i \u00f015\u00de where Is is the pad rotational inertia, i is the tilting angular displacement of the ith pad around its pivot, and M i is the resultant moment of the hydrodynamic forces on the ith pad. Integrating the oil pressure distribution over each pad surface area (see Figure 3(a)) and decomposing it into normal and tangential forces (see Figure 3(b)), we have Fni \u00bc Z a pi cos da, Fti \u00bc Z a pi sin da \u00f016\u00de where Fni is the hydrodynamic force acting perpendicular to the pad surface, Fti is the hydrodynamic force acting tangential to the pad surface, pi is the oil pressure in the gap between the rotor and the ith pad, and is the angle of the curvature of the pad. Thus, the resultant moment of the hydrodynamic forces on ith pad is given by M i \u00bc Fti s \u00f017\u00de where s is the distance between the pad surface and its pivoting point (see Figure 3(a)). Denoting i the angular displacement of the ith pad, the hydrodynamic forces Fx and Fy acting on the shaft can be calculated by projecting the hydrodynamic forces acting on the four bearing pads into the directions x and y. Thus Fx \u00bc P4 i\u00bc1 Fni cos \u2019i \u00fe i\u00f0 \u00de Fti sin \u2019i \u00fe i\u00f0 \u00de , Fy \u00bc P4 i\u00bc1 Fni sin \u2019i \u00fe i\u00f0 \u00de \u00fe Fti cos \u2019i \u00fe i\u00f0 \u00de , 8>>< >>: \u00f018\u00de where \u2019i is the angular position of the ith pad pivot around the bearing (see Figure 3(b)). The hydraulic system consists of a reservoir, two pumps and two servo-valves. One of the pumps supplies the conventional lubrication to the bearing. The other pump feeds the injection system with high pressurized oil. The purpose of the servo-valves is to control the pressure at which oil is injected, through the pad bores, into the bearing gap (see Figure 2(b)). The dynamics of the oil flow through servo-valves I and II can be described by \u20acQVI \u00fe 2 VI !VI _QVI \u00fe !2 VI QVI \u00bc !2 VI KVI uI t\u00f0 \u00de \u20acQVII \u00fe 2 VII "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002616_demped.2011.6063683-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002616_demped.2011.6063683-Figure4-1.png",
+        "caption": "Fig. 4. Non uniform (top) and uniform (bottom) static air-gap eccentricity.",
+        "texts": [
+            "): x l x r r \u03b3\u03b3 =)( (4) After development, the turn function x)(nrk ,\u03b8 and the winding function x)(Nrk ,\u03b8 of the rotor loop k on versus x are given by: \u221e = +\u2212\u2212\u2212+= 1 2 1 cos 2 sin 12 2 , h r rr rk x l )(kh)(h h x)(n r r \u03b3\u03b8\u03b8\u03b8 (5) \u03b8 x r\u03b3 r\u03b1 Using relation (1) and when the skewing effect is considered, the different inductances have similar formulas as presented without skewing effect [13], except that the statorrotor mutual inductances are multiplied by a skewing factor ( ) ske h k for each harmonic component h : ( ) 22 sin rrske h hhk \u03b3\u03b3= (11) ( ) ( ) \u2212\u2212\u2212\u2212+\u2212\u00d7 \u221e = = 23 2 1 2 1 cos 1 2 sin 22 04 0 r ske h tsr p . qr)(krhp h )krhp ( h bhK P N g rl kqh L \u03b3 (12) 2. Uniform and non uniform static eccentricity The air-gap function for a non uniform static eccentricity can be represented by [11,12,14] (Fig. 4): The airgap variation can be expressed as: ( ) ( )\u2212\u2212= \u03b8\u03b4\u03b8 cos11, 0 ex s l x gxg (13) and from the approximation of the first harmonic, one can get: ( ) ( )\u2212+=\u2212 \u03b8\u03b4\u03b8 cos11 1 , 0 1 ex s l x g xg (14) One can notice that the modified winding function x)(sqM ,\u03b8 is different from the winding function x)(sqN ,\u03b8 only in the case of 1=p : ))1p \u03b8\u03b8 (sqN(sqM =\u2260 (15) +\u2212\u2212== 3 2 cosK1))1p 0b \u03c0\u03b8 \u03c0 \u03b4\u03b8\u03b8 q l lN (N(M ex st sqsq (16) The modified rotor winding function x)(M rrk ,,\u03b8\u03b8 can be expressed independently from the value of p as: \u2212\u2212+\u2212\u2212 \u2212\u2212+\u2212\u2212 = 22 1 sin 2 2 s 2 cos 2 s 22 1 cos 2 2 s 2 s 2 1 ),,),, rr r rr rr r r r rexr s r rex s rrkrrk k in in L L k in in L L x(Nx(M \u03b4\u03b1\u03b8\u03b3 \u03b3 \u03b3\u03b1 \u03c0\u03b3 \u03b4 \u03b4\u03b1\u03b8\u03b3 \u03b3 \u03b1 \u03c0 \u03b4 \u03b8\u03b8\u03b8\u03b8 (17) By using (1) one can obtain the different formulas of the stator and rotor inductances and the stator and the stator rotor mutual inductances are dependent on the value of ,p ( 1p = )1or p \u2260 : ( ) ( ) ( ) ( )\u2212\u2212+\u2212= rjrj b rjrj b Healthymrjmrj Ng rl Ng rl LL \u03b8 \u03c0 \u03b8\u03b8 \u03c0 \u03b8 2 2 02 2 0 )( sinsin 12 coscos 12 (18) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( )\u2212++\u2212= rk rk rj rkjrk rk rj rkj b Healthyrjrkrjrk B A Ng rl LL \u03b8 \u03b8 \u03b8 \u03c0 \u03b8\u03b8 \u03b8 \u03b8 \u03c0 \u03b8 2 2 2 2 0 )( sin sin sin sincos cos cos cos 12 (19) ( ) ( ) , 2 cos 2 sin, 2 sin 2 1 11 \u2212=\u0392\u2212=\u0391 skerr exr ssker ex s k L L k L L \u03b3\u03b1 \u03b4 \u03b4\u03b1\u03b4 The case of 1\u2260p : )(HealthymsqLmsqL = , )(2121 HealthyqsqLqsqL = (20) ( ) ( ) ( ) ( )( ) ( )( ) ( )( ) ( )( ) ( )( ) ( )( ) ( )( ) ( )( )\u2212+\u00d7\u2212\u2212+ \u2212+ + \u2212+\u00d7\u2212+= = \u2212+\u2212+ = \u2212+\u2212+ \u221e = 2 1 11 2 1 11 1sin 2 1cos 1 1 2 1cos 2 1 2 cos 1 \u03b5 \u03b5\u03b5 \u03b5 \u03b5 \u03b5 \u03b8\u03b8\u03b3\u03b4 \u03b3 \u03b8\u03b8\u03b4\u03b8 \u03b5\u03b5 \u03b5\u03b5 rkrkq ske hp rsr hp s rex rkrkq ske hp sr hp ex s Healthy rkq ske h sr hprksq hpkhp hp M L l hpkM L l hpkML h (21) The case of 1=p : 2 0 2222 0 )()1( 3 2 cos1 2 +\u2212\u2212== \u03c0\u03b8 \u03b4 q L LK N g rL LL ex bst Healthymsqpmsq (22) ++\u2212\u2212== 3 2 cos 3 2 cos1 2 2010 2222 0 )(21)1(21 \u03c0\u03b8\u03c0\u03b8 \u03b4 qq L LK N g rL LL ex bst Healthyqsqpqsq (23) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( )rkq sker sr s rex rkq ske sr ex s rkrkq ske h rsr h rex s rkrkq ske h sr h ex s h Healthy rkq ske h sr hrksq kM L L kM L L hkhM hl l hkM L L hkML \u03b8\u03b3\u03b4 \u03b3 \u03b8\u03b4 \u03b8\u03b8\u03b3 \u03b3 \u03b4\u03b8\u03b8\u03b4 \u03b8 sin 2 cos 2 1 cos 2 1 2 sin 2 1cos 1 1 2 cos 2 1 2 cos 1111 1111 1 \u2212+\u2212+ ++\u2212+ + ++\u2212+ += ++++ \u221e = (24) 22 1 rr r rk k \u03b3\u03b1\u03b8\u03b8 \u2212\u2212+= , ( )\u2212+\u2212= 3 2 10 \u03c0\u03b8\u03b8\u03b8 qrkrkq , ( ) 2 1 2 1 rr r rkj kj \u03b3\u03b1\u03b8\u03b8 \u2212\u2212++= , ( ) x)rx ( hp bhKN g rl M tsr x = 2 sin04 , )(,)( 21 HealthysqsqLHealthymsqL , )()( , HealthymrjHealthyrkrj LL are the inductances 21 , sqsqLmsqL , mrjrkrj LL , given in the case of symmetrical rotor respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002494_1559-0410.1299-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002494_1559-0410.1299-Figure12-1.png",
+        "caption": "Figure 12: Free body diagram",
+        "texts": [
+            " (3) There exists a unique aim line on a backboard. The aim line is independent of the shooter\u2019s location on the court. (4) The optimal target point can be pinpointed during a training session that employs the pole and aim line. It is the crossing of the pole and the aim line in the shooter\u2019s line of sight. The results presented in this paper can form the basis for future studies aimed at establishing more effective ways of training players how to shoot the bank shot. The following shows that which appeared in Fig. 4. Figure 12 shows the free body diagram of the ball when it makes contact with the backboard. First refer to Fig. 12. sin\u03b8 tan\u03b2 3 cos\u03b8 1 5\u03b3 a R L =   \u2212 + +        (1) 14 DOI: 10.2202/1559-0410.1299 Brought to you by | Dalhousie University Authenticated Download Date | 5/17/15 4:48 PM The important equations are: Equation (2a) follows from conservation of linear momentum in the x direction and assumes a linear visco-elastic collision in which vx and vx\u2019 denote the x components of velocity just before and after contact, and  is the coefficient of restitution (Silverberg and Thrower, 2001). Equation (2b) follows from linear impulse-momentum in the y direction in which m denotes ball mass, Fy denotes the y component of force acting on the ball by the backboard, R is ball radius, and vy and vy\u2019 denote the y components of velocity just before and after contact"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003879_20131120-3-fr-4045.00005-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003879_20131120-3-fr-4045.00005-Figure3-1.png",
+        "caption": "Fig. 3. Guidance algorithm scheme",
+        "texts": [
+            " The cross-track error is then calculated as \u03b5r = |EUAV \u2212mNUAV \u2212 (E1 \u2212mN1)|\u221a m2 + 1 (1) m= E2 \u2212 E1 N2 \u2212N1 (2) (3) Look-ahead or proximity distance, to define the minimum distance of the UAV from the next waypoint to begin turning. When the distance beetwen the aircraft and the next waypoint is less than this distance, the waypoint is reached and the aircraft can move to the next waypoint. Considering these hypotheses and aspects, three main phasesof the guidance sequence have to be analyzed. The first phase is the waypoint approach, i.e. when the UAV gets closer to the waypoint with a theoretical straight flight and gets ready to begin turning around the waypoint. In Fig. 3 this phase is identified by the red dotted line before the point A. The aircraft is flying from the waypoint WP(n\u22121) to the waypointWPn with an assigned speed and altitude. It is assumed the waypoint has been reached when the UAV flies into the imaginary circle centred in the waypoint WPn with radius equal to the defined proximity distance, set to 10 m according to the MH850 dynamic constraints. From this moment the UAV can begin turning around the waypoint in the direction of the next waypoint WP(n+1). In the second phase the aircraft turns around the waypoint. The maneuver is accomplished according with the turn rate constraints function of the cruise speed and of the bank angle. Referring to the Fig. 3, the turn segment is identified by the red dotted curve between the points A and B. It starts when the distance of the vehicle from the waypoint WPn is equal to the proximity distance (point A) and it ends when the longitudinal axis of the vehicle is aligned to the segment WPn,WP(n+1) (point B). The turn is characterized by two parameters: the angle of turn (\u03c8) and the turn radius (Rturn). The first is equal to the difference between the heading angle of the segment WPn,WP(n+1) and the heading angle of the UAV at the point A",
+            " The second depends on the bank angle and on the flight speed and is evaluated as Rturn = V 2 g \u221a n2 \u2212 1 . where V is the aircraft speed, g is the gravity acceleration and n = 1 cos\u03c6 is the turn load factor. At the end of the turn a new reference heading angle is calculated, that is the heading angle of the segment that connects the ending point of the turn (B) to the next waypoint (WP(n+1)). The third segment is related to the straight flight that starts at the end of the last turn and ends at the beginning of the next turn. In Fig. 3 it has been represented by the red spotted line between the points B and C. In this phase the cross-track error control is applied [7]. Moreover, during the straight flight, to avoid continous corrections of the trajectory and of the flight parameters by the elevons, a no corrections region is defined around the waypoint connecting segment. This means that the correction on the heading angle is imposed only when the UAV cross-track error is larger than an assigned value (i.e. maximum acceptable cross-track error)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001697_icece.2010.784-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001697_icece.2010.784-Figure1-1.png",
+        "caption": "Figure 1. Interaction of rotor and stator magnetic forve",
+        "texts": [
+            " The whole starting procedure is coupled with the detection of the rotor position, which can avoid reverse rotation and starting failure. II. INITIAL ROTOR POSTION ESTIMATION In interior PM BLDC, the permanent magnet rotor has salient poles. The saturation of magnetic field depends on the relative location of rotor magnetic force and stator magnetic force, so the equivalent inductance of a stator winding varies due to the saturation effect. 978-0-7695-4031-3/10 $26.00 \u00a9 2010 IEEE DOI 10.1109/iCECE.2010.784 3212 Two kinds of interaction of rotor and stator magnetic force are shown in Fig. 1. One is demagnetization effect with equivalent inductance L1 in Fig. 1(a), the other is enhanced magnetic effect with equivalent inductance L2 in Fig. 1(b). The equivalent inductance will be smaller in a more saturated magnetic field, so L1 <L2. In conclusion, the equivalent inductance of field winding is relative to the rotor position. B. Initial rotor position estimation method In fact, winding inductance can not be measured directly. It can only be obtained by comparing the rate of change of the current flowing through the stator winding. In order to create different equivalent inductances when rotor position is fixed at standstill, voltage vectors with different directions will be applied to the stator windings"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000931_bfb0119383-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000931_bfb0119383-Figure2-1.png",
+        "caption": "Figure 2. (Left) Dual diff drive mode. (Right) Cylinder rolling mode.",
+        "texts": [
+            " Section 4 address previous work, and Section 5 is a discussion and conclusion. 2. T h e M o b i p u l a t o r The Mobipulator (Figure 1) looks like a small car, with four independently powered wheels. None of the wheels are steered. We envision several different modes of manipulation/locomotion, of which five have been demonstrated so far: 9 T r a n s l a t i o n m o d e . This mode is pure locomotion. To translate forward or backward, all four wheels are driven at equal rates. 9 D u a l d i f f d r i v e m o d e . This mode (Figure 2) combines locomotion and manipulation. To maneuver a piece of paper on a desktop, one pair of wheels drives the paper relative to the robot, while the other pair drives the robot relative to the desktop. 9 I n c h w o r m m o d e . This mode (Figure 1) is pure manipulation. With all four wheels on the paper, by quick alternation of the front wheels and rear wheels, the paper can be advanced incrementally beneath the robot. 9 C y l i n d e r ro l l ing m o d e . This mode (Figure 2) is inspired by dung beetles. The robot 's front wheels are placed on a cylinder. It uses its rear wheels to propel itself forward while its front wheels t u rn in the opposite direction to roll the cylinder forward. 9 S c o o t m o d e . With all four wheels on the paper (Figure 1) the robot accelerates hard and then decelerates hard. During the high acceleration the paper slips backward, and during the high deceleration the paper slips forward. There is a net forward motion of the paper. While we have seen all of these techniques work in the lab, so far we have performance data only for dual diff drive mode"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002538_0029-5493(70)90146-9-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002538_0029-5493(70)90146-9-Figure2-1.png",
+        "caption": "Fig. 2. Idealisation of lubrication system.",
+        "texts": [
+            " Denoting by m the normal to the stationary surface directed towards the moving one we find the resulting force vector form R = f m p dS = npi = n a p = l i p , (21) s where n is a (3 \u00d7 su) row super matrix of which a typical sub-matrix is = f m , o d~2 (3 \u00d7 v) . (22) ~2 The integral (22) extends over one dement. 3. Plane surfaces We detail in the present section one finite element formulation for the case when the two surfaces are plane and hence v and m constant vectors. Our analysis presumes triangular TRIM and TRIC-like elements, but it is equally possible to develop a theory for quadrilateral elements. We show a characteristic net in fig. 2 together with a typical TRIC 10 element. The interpolation matrix co for the TRIC family is given in eq. (10) of ref. [1]. A different presentation of co may be found in ref. [2] in this issue, eqs. (19) to (21); see also fig. 1 of this reference. The corresponding set Pl of nodal values is built up from the pressures at the nodal points. For a TRIC 10 element the reader is referred to fig. 2, from which he will deduce PI = { p l P2 \" - -P lO}\" (23) When either of the quantities h, #, ah/Ot and q vary within a single element, the matrix \u00a2o may be used to 228 J.H.ARGYRIS and D.W.SCHARPF interpolate h,/a or 1//~, ~h/at and q from an appropriate set of nodal values. The matrix D x of eq. (8) is established for the TRIC family in ref. [ 1 ]. Using the notation of this last reference we have D x = L R D , (24) where L R and D are given in eqs. (35) and (43) of ref. [1]. The evaluation of the integrals (10) to (13) and (22) can now be carried out immediately, using, in general, a numerical technique"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001888_s12239-010-0070-9-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001888_s12239-010-0070-9-Figure3-1.png",
+        "caption": "Figure 3. Half-car active suspension system.",
+        "texts": [
+            " Finally, we prove that the sphere is a global attractor for the output tracking error of system (11). From (114) and (70), we get (122) For , we have . Thus, any sphere defined by (123) is a global final attractor for the tracking error system of the nonlinear control systems (3.1). Furthermore, it is easy to see that, for , we have that is, According to the comparison theorem, we get Therefore, Consequently, we obtain , i.e., the convergence rate toward the sphere is equal to \u03b1*/ 2. This completes our proof. The model of a half-car suspension system is shown in Figure 3 (Karnopp, 1983; Krtolica and Hrovat, 1990, 1992; Campos et al., 1996). The half-vehicle suspension model is represented as a nonlinear four-degree-of-freedom (DOF) system. It consists of a single spring mass (car body) connected to two non-spring masses (front and rear wheels) at each corner. The spring mass is free to heave and pitch, while the non-spring masses are free to bounce vertically Br ytotal r> L \u00b7 0< Br (124) with respect to the spring mass. The suspensions between the spring mass and non-spring masses are modeled as linear viscous dampers and spring elements, while the tires are modeled as simple linear springs without damping. From Figure 3, it can be seen that the displacements of the sprung mass are given by Front wheel (125) Rear wheel (126) where zsf is the front body displacement, zsr is the rear body displacement, a is the distance between front axle and center of gravity, b is the distance between the rear axle and center of gravity, \u03b8 is the pitch angle, and z is the displacement of the center of gravity. Equivalent forces in both wheels are given by Front wheel (127) Rear wheel (128) where and are the front and rear force inputs; Bf and Br are the front and rear damping coefficients; kf and kr are the front and rear spring coefficients; and zuf and zur are the front and rear wheel displacements"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003709_012082-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003709_012082-Figure1-1.png",
+        "caption": "Figure 1. UNSW gearbox test rig.",
+        "texts": [],
+        "surrounding_texts": [
+            "The dynamic simulation of the gearbox test rig was carried out based on a 34 DOF model developed earlier [1]. The model accounts for time varying nonlinearity of gear and bearing stiffness and random slippage in the bearings. The model is capable of simulating bearing faults (both localised and extended faults in the inner and outer race) in addition to spalls and cracks in the gears. Only two degrees of freedom were used to represent the casing, a low frequency rigid body mode and a high frequency resonance at 15 kHz excited by the bearing faults. The LPM model was found to be adequate to simulate localised bearing faults [1]. However, the limited representation of the casing resulted in the poor spectral matching over a wide frequency range and lack of interaction with the bearing faults (experienced in practice) mainly in case of extended faults. The spectral matching was improved considerably by applying the generated forces to the casing FRFs."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.5-1.png",
+        "caption": "Fig. 1.5. Wheel suspension and steering compliance resulting in additional steer angle qJl.",
+        "texts": [
+            " In the next subsection, this non-linear effect will be incorporated in the effective axle characteristic. Effective Non-Linear Axle Characteristics To illustrate the method of effective axle characteristics we will first discuss the determination of the effective characteristic of a front axle showing steering compliance. The steering wheel is held fixed. Due to tyre side forces and selfaligning torques (left and right) distortions will arise resulting in an incremental steer angle ~ of the front wheels (~c~ will be negative in Fig. 1.5 for the case of just steer compliance). Since load transfer is not considered in this example, the situation at the left and right wheels are identical (initial toe and camber angles being disregarded). The front tyre slip angle is denoted with a~. The 'virtual' slip angle of the axle is denoted with aal and equals (cf. Fig. 1.5): O~al - 0~1 - ~ffcl (1.24) where both a~ and ~c~ are related with Fy~ and Mz~. The subscipt 1 refers to the front axle and thus to the pair of tyres. Consequently, Fy~ and Mz~ denote the sum of the left and right tyre side forces and moments. The objective is, to find the TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 13 function Fyl(aal ) which is the effective front axle characteristic. Figure 1.6 shows a graphical approach. According to Eq.(1.24) the points on the Fy~(al) curve must be shifted horizontally over a length ~cl to obtain the sought Fyl(a~l)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003880_s00521-012-1326-2-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003880_s00521-012-1326-2-Figure3-1.png",
+        "caption": "Fig. 3 The double pendulums system [29]",
+        "texts": [
+            " Then, the asymptotically convergence of tracking error is guaranteed. All the signals of the resulting closed-loop system are globally uniformly ultimately bounded. h Remark 2 Theorem 1 introduces the adaptive backstepping controller design for system (1) with second order. This approach can be extended to the nth order system easily by repeating the design of virtual controllers. Example 1 Double Pendulums System in Nonlinear Nonaffine Form Consider the tracking control of two degree-of-freedom double pendulums [29] (shown in Fig. 3). The two rods rotate in the vertical plane and two connecting joints are derived by torque control. All frictional forces are ignored here. The following motion equations can be derived [29] M1\u00f0t\u00de M2\u00f0t\u00de \u00bc 1 3 l2 1\u00f0m1 \u00fe 3m2\u00de\u20ach\u00fe 1 2 gl1\u00f0m1 \u00fe 2m2\u00de sin\u00f0h\u00de\u00f0t\u00de \u00fe 1 2 l1l2m2 cos\u00f0/\u00f0t\u00de h\u00f0t\u00de\u00de\u20ac/\u00f0t\u00de 1 2 l1l2m2 _/2\u00f0t\u00de sin\u00f0/\u00f0t\u00de h\u00f0t\u00de\u00de M2\u00f0t\u00de \u00bc 1 2 l1l2m2 cos\u00f0/\u00f0t\u00de h\u00f0t\u00de\u00de\u20ach\u00f0t\u00de \u00fe 1 3 m2l2 2 \u20ac/\u00f0t\u00de \u00fe 1 2 gl2m2 sin\u00f0/\u00f0t\u00de\u00de 1 2 l1l2m2 _h2\u00f0t\u00de sin\u00f0/\u00f0t\u00de h\u00f0t\u00de\u00de: \u00f034\u00de where h(t) and /(t) are defined in Fig. 3, and M1\u00f0t\u00de; M2\u00f0t\u00de are the torques acting on the connecting joints of the rods 1 and rods 2. We denote the functions fij, i = 1, 2; j = 1, 2, 3, as follows. f11\u00f0h;/\u00de \u00bc 12 l2 1\u00bd4m1 \u00fe 12m2 9m2 cos2\u00f0/ h\u00de ; f12\u00f0h;/\u00de \u00bc 12l2 \u00fe 18l1 cos\u00f0/ h\u00de l2 1l2\u00bd9m2 cos2\u00f0\u00f0/ h\u00de\u00de 4m1 12m2 ; f21\u00f0h;/\u00de \u00bc 18 cos\u00f0/ h\u00de l1l2\u00bd9m2 cos2\u00f0/ h\u00de 4m1 12m2 ; f22\u00f0h;/\u00de \u00bc 12m1 \u00fe 36m2 l2 2m2\u00bd4m1 \u00fe 12m2 9m2 cos2\u00f0/ h\u00de \u00fe 18l2 cos\u00f0/ h\u00de l1l22\u00bd4m1 \u00fe 12m2 9m2 cos2\u00f0/ h\u00de ; f13\u00f0h; _h;/; _/\u00de \u00bc 9gm2 sin\u00f02/ h\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de 9l1m2 _h2 sin\u00f02/ 2h\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de \u00fe 12l2m2 _/2 sin\u00f0/ h\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de \u00f015gm2 \u00fe 12gm1\u00de sin\u00f0h\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de ; f23\u00f0h; _h;/; _/\u00de \u00bc 12l2 1 _h2 sin\u00f0/ h\u00de\u00f0m1 \u00fe 3m2\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de 9l2m2 _/2 sin\u00f02/ 2h\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de 9g sin\u00f0/ 2h\u00de\u00f0m1 \u00fe 2m2\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de \u00fe 3g sin\u00f0/\u00de\u00f0m1 \u00fe 6m2\u00de l1\u00bd15m2 \u00fe 8m1 9m2 cos\u00f02/ 2h\u00de : \u00f035\u00de Then, Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002482_j.triboint.2009.12.042-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002482_j.triboint.2009.12.042-Figure5-1.png",
+        "caption": "Fig. 5. Partial parameters for a hill region.",
+        "texts": [
+            " This means, that the hill regions, where the vertical difference between the peak and its nearest saddle point is smaller than 30 nm, are defined as insignificant structures. The feature based parameters has also been calculated by setting the threshold at 20 and 40 nm. The results show the same relationship between the parameters and the normal loads like by using a threshold at 30 nm. As Fig. 4b shows, each hill region is bordered by its course line and has a shape that is affected by the solid contact during sliding. In order to characterize the shapes of hill regions, the following feature based parameters are defined. One hill region is shown in Fig. 5. The peak height (Sph) is defined as the average difference between a peak and its course line. The hill area (Sha) is the projected area of a hill region. The aspect ratio (Sar) denotes the aspect ratio of a hill area in the sliding direction and is defined as the root of quotient of the sums of the distance square a2 i and b2 i [10] (see Fig. 6): S2 ar \u00bc Pn i \u00bc 1 a2 iPn i \u00bc 1 b2 i \u00f01\u00de If the aspect ratio is small, it means the structure tends to have a longish form. Contrariwise, the structure tends to have a round form",
+            " In order to obtain the curvature of the fitted polynomial surface in the desired direction, the data is transformed into a new coordinate system (v,w) in which the direction of the w axis is parallel to the sliding direction (Fig. 8). The polynomial surface is defined as: f \u00f0v;w\u00de \u00bc a00\u00fea10v\u00fea01w\u00fea20v2\u00fea11vw\u00fea02w2 \u00f02\u00de Setting w to 0, the Eq. (2) can be formed as f \u00f0v\u00de \u00bc a00\u00fea10v\u00fea20v2 \u00f03\u00de The second derivative of f \u00f0v\u00de is defined as the peak curvature in the v direction: f 00 \u00f0v\u00de \u00bc j2a20j \u00f04\u00de Similarly, the peak curvature in the w direction is f 00 \u00f0w\u00de \u00bc j2a02j \u00f05\u00de The peak root mean square deviation (Spq) denotes the root mean square deviation of the absolute peak height (see Fig. 5) in a sampling area. This parameter indicates the vertical distribution of peaks. The peak density (Spd) is the number of the peaks in a unit area. This parameter describes the lateral distribution of peaks. The parameters (Sph, Sha, Sar, Spcs, Spcps) described above indicate the geometrical information of a single hill region. In order to represent the characteristic of the hill regions in a sampling area, the mean value of these parameters (mSph, mSha, mSar, mSpcs, mSpcps) is used. Even though the Vickers hardness of the ceramic ball is about twice as large as the hardness of the steel disc, the wear of the ceramic ball is not negligible"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002694_icra.2012.6224600-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002694_icra.2012.6224600-Figure3-1.png",
+        "caption": "Fig. 3. Simplification to the outward recursion that is provided by the acceleration propagator k1X 1. It\u2019s transpose, the force propagator 1XT k1 was calculated during the inward recursion. As a result, three recursive steps for the calculation of \u039b\u22121 k1 k2 are able to be replaced with one matrix multiplication.",
+        "texts": [
+            " For each k \u2208 ES(i) we have the recursion: \u039b\u22121i k := Li iXp(i) \u039b\u22121p(i) k +Ki kX T i , (13) while \u039b\u22121i k := Li iXp(i) \u039b\u22121p(i) k for each k /\u2208 ES(i). Thus, if i does not support end-effector k2, the acceleration influence of fe k2 propagates from p(i) to i according to an articulated acceleration transform (as if the joint i were free to move). As a result, for any other end-effector k1 in the subtree at i we have the following: \u039b\u22121k1 k2 := k1X p(i) \u039b\u22121p(i) k2 . (14) This simplification is illustrated in Fig. 3 for a basic example. We note that this simplification is first possible when i is the child of ancest(k1, k2). In the example figure, an acceleration propagator may be applied to calculate \u039b\u22121k1 k2 recursively through links 1, 2, and 3. Yet, computational savings can occur when \u039b\u22121k1 k2 is computed directly at the common ancestor of k1 and k2 through the use of k1X 1. With this insight we define the following set: GCA(i) = {(k1, k2) | k1 < k2 and i = ancest(k1, k2)} (15) which contains all end-effector pairs that have a greatest common ancestor at i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003933_ijcmsse.2011.042825-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003933_ijcmsse.2011.042825-Figure1-1.png",
+        "caption": "Figure 1 Configuration of the bearing system",
+        "texts": [
+            " Here, it is sought to analyse the performance of a magnetic fluid-based squeeze film behaviour between curved transversely rough rotating circular plates, where in, the upper plate lies along a surface determined by a hyperbolic function while the lower plate is along the surface governed by a secant function. These types of plates have been chosen purely to increase the load carrying capacity, in turn, leading to an overall improved performance of the bearing system. In addition to this it may provide an additional degree of freedom, which can be seen through the forms associated with different curved surfaces. The geometrical configuration of the rotating bearing system is displayed in Figure 1. The bearing surfaces are assumed to be transversely rough. Following Christensen and Tonder (1969a, 1969b, 1970) the thickness h(x) of the lubricant film is considered as ( ) ( ) sh x h x h= + (1) where ( )h x is the mean film thickness while hs is the deviation from the mean film thickness characterising the random roughness of the bearing surfaces. The deviation hs is assumed to be stochastic in nature and described by the probability density function ( ) ,s sf h c h c\u2212 \u2264 \u2264 (2) where c is the maximum deviation from the mean film thickness"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002929_s1068798x11010278-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002929_s1068798x11010278-Figure2-1.png",
+        "caption": "Fig. 2.",
+        "texts": [
+            " We also assume that the elastic and dissipative effects are proportional to the corresponding generalized coordinates and velocities [3]. The distances at which the elastic and dissipative forces act are constant. The system\u2019s kinetic energy consists of the kinetic energy of the operational system T1 and the kinetic energy of the unbalanced mass T2: T = T1 + T2. The kinetic energy of the operational system is (2) The kinetic energy of the unbalanced mass is (3) To determine the center of gravity coordinates of the unbalanced mass, we consider Fig. 2, where O0 is the center of gravity of the operational system at time t0 = 0; O is the center of gravity of the operational sys tem with coordinates x and y at any time t. The center of gravity of the unbalanced mass has the coordinates \u2013(h + r), 0 at time t0 = 0 and xg, yg at time t = t1. T1 m1 2 x\u00b7 2 y\u00b7 2+( ) 1 2 J1\u03c8\u00b7 2 .+= T2 m0 2 x\u00b7 g 2 y\u00b7 g 2+( ) 1 2 J0\u03c92 .+= \u03be \u03c8 x \u03b7 yO0 by cy bx O h p l1l1 cy by p 2 bx cx cx S m0 \u03c9 t \u2013 90\u00b0 pS \u03c9 A1A0 r 1 2 3 4 5 6 7 896 5 7 8 4 9 Fig. 1. RUSSIAN ENGINEERING RESEARCH Vol. 31 No. 1 2011 MECHANICS OF AN ABRASIVE MEDIUM IN A VIBRATING CONTAINER 17 From Fig. 2, taking into account that O0A0 = OA1 = x0, we find that Substituting and into Eq. (3), we obtain (4) The system\u2019s potential energy consists of the energy of the left and right shock absorbers For the system (5) Taking account of Eqs. (2), (4), and (5), we write the Lagrange function Substituting the derivatives u (k = 1, 2, 3) into Eq. (1), we obtain the required differ ential vibrational equations of the system (6) where Qx, Qy, Q\u03c8 are the generalized forces, which, for present purposes, are dissipative forces"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-149-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-149-1.png",
+        "caption": "Figure 3-149: Severe fretting on the female splines. Note the even depth and the diminishing pattern following the trapped male spline segment (magnified 0.5 ).",
+        "texts": [],
+        "surrounding_texts": [
+            "In all machinery component failure analysis activities, it is necessary to understand the design and functional peculiarities of the component and of the machinery system it is associated with. This is especially true when analyzing gear-coupling failures or distress signs. The following example will prove this point. In 1972, Petronta had to shut down their 8500-hp propylene compressor after experiencing high lateral vibrations on the compressor. The compressor train configuration is similar to that in Figure 9-21. Upon inspection of the high-speed coupling, it was discovered that large portions of the hub gear teeth were broken. The coupling, a spacer type B, Figure 3-146, had been in service for 15 years and had received inspections combined with regreasing operations at two-year intervals. Because the failure was judged extremely serious, it was decided to have the failed coupling analyzed by a reputable engineering laboratory. Their failure analysis report read as follows: 1.0 Introduction (narrative deleted) 2.0 Investigation Procedures and Results 2.1 Visual Examination 2.1.1 The submission, as received, consisted of a male and female splined component and numerous spline segments which had separated from the male component. Some segments were jammed in the bottom of the female spline, necessitating trepanning out of a retainer insert to facilitate their removal. 2.1.2 All surfaces were lightly rusted and dry (i.e. no trace of grease) and a dry, finely divided, red rust-like deposit was present at the bottom of the female spline. With two exceptions, the fracture faces on the male spline segments left integral with the hub were either in a circumferential plane or at an acute angle with the spline end faces. Where circumferential fracture faces were observed, most, and in some cases all, of the entire spline was absent and the surfaces were generally scalloped and heavily fretted (Figure 3-148). In all cases the scalloping had undercut the spline end-face radii at the bases of the individual teeth. The acute angled fracture faces were predominantly brittle, and characteristic \u201csunburst\u201d brittle fracture patterns were seen originating at the loaded flanks on the root radii of several splines. Occasional chevron marks in the fracture surfaces indicated that most of the cracks had propagated in the direction opposite to rotation. 2.1.3 The broken pieces of male splines had fracture surface markings which corresponded with neighboring pieces or with the fracture surfaces of male splines left integral with the hub. However, in two instances beach-like markings were observed on pieces where the fracture front was emerging at the unsupported radial/circumferential spline end-face surface at an acute angle. These pieces did not have common surfaces with the integral male splines. Many of the pieces had \u201csunburst\u201d fracture patterns focusing at the loaded spline flank root radii. 2.1.4 Examination of the female component of the coupling indicated that none of the splines were fractured or visibly cracked, but heavy fretting was observed on the load-bearing Machinery Component Failure Analysis 213 2.1.5 Upon reassembly of all the pieces of the male coupling, it was apparent that the indi- vidual male splines were fretted to different degrees in much the same manner as their"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001398_02286203.2007.11442442-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001398_02286203.2007.11442442-Figure2-1.png",
+        "caption": "Figure 2. Body model segment shape showing principal coordinate system.",
+        "texts": [
+            " The golfer was modelled as a variable full-body, multilink, three-dimensional humanoid mechanism made up of 15 rigid segments interconnected with joints. The model contains 15 body segments: head, neck, thorax, lumbar, pelvic, upper arm (2), forearm (2), thigh (2), lower leg (2), and foot (2). All segments are defined by their adjacent joints with exceptions of the neck (C1\u2013C8), thorax (T1\u2013 T12), and lumbar (L1\u2013L5 and S1\u2013S5) which are defined by the associated vertebrae. The individual body segments are ellipsoid in shape (see Fig. 2) with the segment size, mass, and inertia properties determined from gender, age, and overall body height and weight, or from local segment measurements using the GeBod database accessible through the ADAMS software, or from one of the many references concerning body segment mass properties [10]. The body segment reference coordinate systems, established when the subject is standing in the standard anatomical position, place the Z-axis pointing downward with the exception of the feet which point forward parallel to the long axis of the foot segment"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure10-1.png",
+        "caption": "Fig. 10 Line of action of external gear",
+        "texts": [
+            " 9, in a way resembling the load distribution on the gear surface, and then gear teeth are cut with gear teeth cutting methods. In the injection method, the gear is obtained with the help of dies, as explained before. 3.4 Hertzian stress along contact path Gear teeth are subjected to Hertz contact stresses and various surface damages. Overloading is one of the basic factors that cause abrasion, and scoring. Generally, a good correlation has been observed between spur gear surface fatigue failure and the computed elastic surface stress (Hertz stress) [20]. As shown in Fig. 10, the radius of the pinion at the initial contact point A is \u03c11 and that of the gear is \u03c12. For external spur gears, the Hertzian contact stress at point A is \u03c3H = \u221a \u221a \u221a \u221a \u221a \u221a F ( 1 \u03c11 + 1 \u03c12 ) \u03c0b ( 1\u2212\u03bd2 1 E1 + 1\u2212\u03bd2 2 E2 ) (1) where F is normal force, b is the gear tooth face width, E is Young\u2019s modulus, and \u03bd is the Poisson ratio. One of the fundamental relationships is evident from this equation. Because of the increased contact area with load, stress increases only as the square root of load F (or square root of load per mm of face width, F/b)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure9.3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure9.3-1.png",
+        "caption": "Fig. 9.3-1 Displacement of a longitudinal vibration rod.",
+        "texts": [
+            "2-4 Initial displacement of a string. VibrationAnalysis_txt.indb 314 11/24/10 11:49:35 AM Continuous Systems | Chapter 9 315 Next, for zero initial velocity, Cn = 0 in Eq. c. Hence, Eq. 9.2-24 becomes u x t H ab n n a n x n ct n ( , ) sin sin cos= = \u221e \u22112 12 2 2 1 L L L L (f) Particularly, if a=b=L/2, Eq. f then becomes [9.2] u x t H n n xn n ( , ) ( ) ( ) sin ( ) cos (= \u2212 \u2212 \u2212\u2212 = \u221e \u22118 1 2 1 2 1 2 2 1 2 1 L n ct\u22121) L (g) 9.3 Longitudinal Vibration of Rods Consider a thin homogeneous rod, as represented in Fig. 9.3-1 (A) [9.2, 9.3]. Consider a diff erential element with length dx, as shown. Let u(x,t) be the axial, or longitudinal, displacement of the element. Fig. 9.3-1 (B) provides a force representation for the element where P is the axial force. From Hooke\u2019s law, as a stress-strain relation, we have P AE u x = \u2202 \u2202 , (9.3-1) where A is the cross-section area of the rod, E is the Young\u2019s modulus of elasticity, and \u2202 \u2202u x/ is the strain. Applying Newton\u2019s law with Fig. 9.3-1 (B), we have dx u t P P x dx P( )\u2202 \u2202 = + \u2202 \u2202 \u239b \u239d\u239c \u239e \u23a0\u239f \u2212 2 2 , (9.3-2) where is the mass of the rod per unit length. Substituting from Eq. 9.3-1 into Eq. 9.3-2 and simplifying, we obtain the equation of motion \u2202 \u2202 = \u2202 \u2202 \u2202 \u2202 \u239b \u239d\u239c \u239e \u23a0\u239f 2 2 u t x AE u x (9.3-3) VibrationAnalysis_txt.indb 315 11/24/10 11:49:36 AM Principles of Vibration Analysis 316 If AE is a constant, we have \u2202 \u2202 = \u2202 \u2202 2 2 2 2 2 u t c u x , (9.3-4) where c AE= / is the velocity of propagation of the displacement, or the stress wave, in the rod",
+            " Th us, the solution has the form u x t D n x n c tn n ( , ) sin cos= = \u221e \u2211 L L1 (f) Next, from Hooke\u2019s law, the stress-strain relation is P A E u x x 0 2 0= ( , ) or u x P x EA ( , ) ,0 2 0= (0 < x < L/2) (g) P A E u x x 0 2 0= \u2212 ( , ) L or u x P EA x( , ) ( ),0 2 0= \u2212L (L/2 < x < L) (h) Applying these initial conditions in Eq. f, we obtain D P x AE n x dx P EA x n n L L L = + \u2212\u222b \u222b 2 2 2 2 0 0 2 0 2 L L L L / / sin ( )sin x dx P L AE n n L = 2 1 2 0 2 2 sin (n=1,2,\u2026) (i) Th erefore, the longitudinal vibration of the bar is u x t P AE n n xn n ( , ) ( ) ( ) sin ( ) c= \u2212 \u2212 \u2212\u2212 = \u221e \u22112 1 2 1 2 10 2 1 2 1 L L os ( )2 1n ct\u2212 L (j) VibrationAnalysis_txt.indb 317 11/24/10 11:49:40 AM Principles of Vibration Analysis 318 Example 9.3-2 [9.4] Fig. 9.3-2 shows a test setup for obtaining the frequency response function (FRF) of a structure in an area where the impact hammer cannot reach. Th at is, a uniform bar B is attached to a structure S, where the FRF is diffi cult to be measured by hammer impact testing. A transducer T is mounted on the top of the bar. With a hammer H, the driving point FRF of the top of the bar can easily be measured. Once this FRF is obtained, the FRF of the structure on the bottom of the bar can accurately be calculated based on the driving point FRF of the top of the bar and the physical properties of the bar as explained below. Fig. 9.3-3 shows a free-body diagram of the bar. Th e solution of the wave equation for the bar has the form: u x t C x c C x c t( , ) sin cos sin( ),= +\u23a1 \u23a3\u23a2 \u23a4 \u23a6\u23a5 +1 2 (a) where C1, C2, and are constants that are dependent on the boundary conditions and initial conditions. For this case, the boundary conditions for the left end of the bar are a u C t x1 2 0 2 2= \u2212 =\u2212 + = sin( ) (b) F AE u x AE c C t x 1 0 1= \u2212 \u2202 \u2202 = \u2212 + = sin( ), (c) where a1 is the acceleration of the bar at the left end, and where A is the rod cross-section area",
+            " b to e, the fi nal equations relating the state vectors at the left and right ends are found to be a a F r 2 1 1= +cos sin m (i) F a Fr 2 1 1= \u2212 + m sin cos (j) Let us now defi ne the FRFs as H a F H a F22 2 2 11 1 1 = =and (k) By substituting from Eqs. i and j, H22 is seen to be H H H r r 22 11 11 = + \u2212 sin cos cos sin m m (l) By inspection of Eq. l, we see that the expression is an algorithm for determining the accelerance a/F of the structure as it would be measured in the recessed space (see Fig. 9.3-2). Specifi cally, since it is physically diffi cult to VibrationAnalysis_txt.indb 319 11/24/10 11:49:43 AM Principles of Vibration Analysis 320 take measurements in the recessed space, we can simply attach a slender bar to the desired measurement location and then measure the accelerance H11 at the free end of the bar. Th en, using Eq. l, once H11 is known, the accelerance H22 of the structure is readily determined. 9.4 Torsional Vibration of Rods Consider next the twisting oscillation of a circular rod, as represented in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003842_0954406211400691-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003842_0954406211400691-Figure5-1.png",
+        "caption": "Fig. 5 Training and testing results of the RVM model",
+        "texts": [
+            " From this point of view, RVM has better upgradeability than SVM. In order to further investigate the characteristics of the RVM model, the experimental results of the use of 300 training and 300 testing samples are plotted in the 2D feature space, as shown in Figs 5 and 6. It is seen that the boundaries of the two methods are similar, but the RVM only utilizes three RVs to construct the boundary line while almost all the input vectors near the boundary are chosen as SVs for the SVM. Moreover, as can be seen in Fig. 5, the RVM model provides a full posterior probability distribution (PPD) of class membership rather than make a simple \u2018hard\u2019 binary decision as the SVM. The red line is the classification boundary where the posterior probability equals 50 per cent. The gradient colour from yellow to green represents the PPD where all the points on each of the contour lines have the same posterior probability as shown by text (e.g. 0.8, 0.9). Only the posterior probability of the abnormal class is presented as given by equation (19), and the posterior probability of the normal class can be simply obtained through equation (20)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002693_iecon.2010.5675183-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002693_iecon.2010.5675183-Figure2-1.png",
+        "caption": "Fig. 2. Dynamics of DOF Helicopter",
+        "texts": [
+            " The motors correspond to each one of the actuators of the propellers. The pitch and yaw motors voltage are \u00b124V and \u00b115V respectively. Some others specifications like masses, torques, inertias, and others, are presented in [6]. In order to obtain the mathematical model of the system, A 978-1-4244-5226-2/10/$26.00 \u00a92010 IEEE 162 the angle \u03b8 in the pitch axis and the angle \u03c8 in the yaw axis represent two DOF. The pitch axis is positive when the nose of the helicopter goes up, the yaw axis is positive for a clockwise rotation. Also in the Figure 2, there are thrust forces Fp and Fy for everyone for their respective axis, the torque of pitch is being applied at a distance rp from the pitch axis and a yaw torque is being applied at a distance ry from the yaw axis. The gravitational force Fg pulls down on the helicopter nose. The center of mass is at a distance of l cm from the pitch axis along the helicopter body length. The center of mass of the aircraft, after the transformation of the coordinates, using the pitch and yaw rotation matrices, is given by: \u03b8 \u03b8\u03c8 \u03b8\u03c8 sin cossin coscos cmcm cmcm cmcm lz ly lx = \u2212= = (1) Where lcm is the distance between the center of mass and the intersection of the pitch and yaw axes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003089_acc.2011.5990634-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003089_acc.2011.5990634-Figure1-1.png",
+        "caption": "Fig. 1. The helicopter\u2019s body-fixed frame, the Tip-Path-Plane angles and linear/angular velocity components.",
+        "texts": [
+            " These experimental applications indicate that the linear model proposed by [6] provides a generalized and physically meaningful solution for developing practical linear models for small-scale helicopters. The helicopter motion variables are expressed with respect to a body-fixed reference frame defined as FB = {OB,~iB,~jB, ~kB}, where the center OB is located at the Center of Gravity (CG) of the helicopter. The directions of the body-fixed frame orthonormal vectors {~iB,~jB, ~kB} are shown in Fig. 1. The helicopter\u2019s linear and angular velocity vectors, with respect to the body-fixed frame, are denoted by vB = [u v w]T and \u03c9B = [p q r]T , respectively. The helicopter attitude is expressed by the roll (\u03c6), pitch (\u03b8) and yaw (\u03c8) angles. The helicopter motion variables are shown in Fig. 1. The control input is defined as uc = [ulon ulat ucol uped] T where ucol and uped are the collective controls of the main and tail rotor, respectively. The collective commands control the magnitude of the main and tail rotor thrust. The other two control commands ulon, ulat are the cyclic controls of the helicopter which control the inclination of the Tip-PathPlane (TPP) on the longitudinal and lateral direction. The TPP is the plane in which the tips of the blades lie. The TPP is characterized by two angles, a and b which represent the tilt of the TPP at the longitudinal and lateral axis respectively. The inclination of the TPP can be seen in Fig. 1. The TPP is itself a dynamic system. The adopted linear model represents the dynamic response of the helicopter perturbed state vector from the reference flight condition. In this case, the reference operating condition is hover. The liner state space model is described by: x\u0307 = Ax+Buc (1) where the state vector is given by: x = [u v \u03b8 \u03c6 q p a b w r \u03c8]T The entries of the matrices A and B are given in Table I. These entries are also called stability and control derivatives, respectively. The term g denotes the gravitational constant while \u03c4f is the main rotor\u2019s time constant"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002744_gt2011-46018-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002744_gt2011-46018-Figure5-1.png",
+        "caption": "Figure 5 Curved beam element [17]",
+        "texts": [
+            " But actually, the top foil in the bearing house is curved. So in this paper, the 1-D curved beam model is given to compare with other top foil models. Fig. 4 depicts the one dimensional model of the top foil. One end of the top foil is fixed with the transverse deflection and rotation equal to zero, while the other end is free. The bump strip layer beneath the top foil is in the same situation. The freedom degrees of 1-D model are shown as transverse deflections (wd) and rotations ( z ). The curved beam model is shown in Fig. 5. xc, yc, zc are the natural (local) curved element coordinates, , ,c c cx y z are the generalized element coordinates, and Xc, Yc, Zc are the global coordinates. For thin beams without shear deformation effects and rotary inertia, the radius of neutral axis of bar rn is equal to radius of centroidal line of bar Rc. Pilkey [17] details the elasticity equations for deformation in a thin curved beam. For the beam model, the finite element method is used to obtain the transverse deflections (wd). Curved beam element stiffness matrix k i in the natural (local) curved element coordinate system is: where Q and S are the coefficient matrices for curved beam element and detailed in ANNEX A"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001538_1.3548167-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001538_1.3548167-Figure4-1.png",
+        "caption": "FIG. 4. \u2014 (a) Schematic of a peel test specimen with force applied across the interface; (b) Experimental setup showing a peel test specimen.",
+        "texts": [
+            " The tape was used to separate a third of each of the two sheets to allow the two free ends to be gripped during testing. This combination was compression molded with the same curing conditions (160 \u00b0C, 15 min). A flat interface was thus created between the two sheets as shown in Figure 3 (stage 2). The sheet representing granulates was double cured (A in Figure 3) and the sheet (B in Figure 3) representing the matrix was single cured. In order to quantify the interfacial strength, five separate strips were cut out and as shown in Figure 4, the gripped free ends were pulled apart at a constant rate to separate the strips across the interface. Quantification of interfacial strength between matrix and granulates in this manner has been done in the past.11 The peel energy, P, was calculated25 using: (2) where F is the average peel force; w denotes the width of the peel strips, ~30mm; t is the thickness of each sheet, ~2 mm; W is the elastic strain energy density in the stretched strips. P F w tW= \u2212 2 \u03bb In order to obtain interface between granulates, a similar procedure (as described above) was adopted to prepare the specimens and is shown in Figure 5"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002330_jjap.50.05ed03-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002330_jjap.50.05ed03-Figure2-1.png",
+        "caption": "Fig. 2. (Color online) Schematic diagram of the X-ray microdiffraction optics at BL13XU in SPring-8.",
+        "texts": [
+            " The typical fabrication process of the WOW structure has been described in detail in the previous paper.4) Figures 1(a) and 1(b) show plan and cross-sectional view schematic diagrams of a part of the sample structure used in this study, respectively. The thickness of the thinned Si layer is 10 m. The diameter and the pitch of the TSV interconnect are 10 and 40 m, respectively. The thin Si layer and Si wafer were bonded by using CYCLOTENE resin as an adhesive material. The X-ray microdiffraction measurement was performed at the beamline BL13XU in the super photon ring 8GeV (SPring-8).12) Figure 2 shows a schematic diagram of the X-ray microdiffraction optics at BL13XU. A synchrotron radiation light from the storage ring was monochromatized to an energy of 10 keV ( \u00bc 0:123984 nm) by using the Si(111) double-crystal monochromator and was focused on a sample by using a zone plate. The exposed area of an incident X-ray microbeam on a sample was estimated to be 0:76 0:56 m2. High-speed data acquisition was also realized with a charge coupled device (CCD) image sensor. The position of the incident X-ray microbeam was controlled by checking the CCD microscope"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002691_icca.2010.5524420-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002691_icca.2010.5524420-Figure2-1.png",
+        "caption": "Fig. 2 Structure of the tri-rotor aircraft",
+        "texts": [
+            " Some kinds of tri-rotor aircraft [8-11] were designed previously, at the same time, the corresponding controller was proposed to stabilize them. All those aircraft have the ability of hovering, however, they did not have the ability of high-speed forward flight. If fixed wings can be added into tri-rotor aircraft, then the aircraft not only have the ability of VTOL, but also have the ability of high-speed forward flight. On the other hand, mode transition control from hovering to forward flight and reverse is difficult. In this paper, a tri-rotor aircraft with fixed wings is designed (see Fig. 2), and the ability of VTOL and high-speed forward flight can be realized. Moreover, a PD controller based on back-stepping technique is designed to achieve mode transition control of the aircraft. II. DESIGN OF TRI-ROTOR AIRCRAFT In order to make the aircraft have the ability of VTOL and forward flight with high speed, a tri-rotor aircraft is designed as follows. This work has been supported by National Natural Science Foundation of China (60774008). The authors are with National Key Laboratory of Science and Technology on Holistic Control, Department of Automatic Control, Beijing University of Aeronautics and Astronautics, Beijing 100191, China. (Email: fphbuaa@126.com, wangxinhua04@gmail.com, kycai@buaa.edu.cn) As it is shown in Fig. 2, the aircraft has three rotors and two commutating devices which are respectively driven by motors and small actuators. From components 4-8, it is shown that the three ro-tors rotate synchronously with the same angle \u03b2 (shown in Fig. 5). Firstly, actuator 4 drives the driving gear 5, then, the driving gear 5 drives the transmission shaft 6, finally, the transmission shaft 6 drives the front driven gear 8 and the tail driven gear 7. So the three rotors can rotate synchronously with the same angle \u03b2 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001857_10402000903283292-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001857_10402000903283292-Figure9-1.png",
+        "caption": "Fig. 9\u2014Simplified lip geometry: (a) grease volume in the vicinity of the lip contributing to the lift, and (b) grease volume in the contact zone with contact width b and film thickness h(x).",
+        "texts": [
+            " The second normal stress might play a role here, but this has not been investigated. In the contact zone the shear rates are extremely high due to the small film thickness. Film thickness measurements on oillubricated seals show an oil film thickness of 1-5.5 \u00b5m (Van Leeuwen and Wolfert (14)). Thinner films are expected for grease-lubricated seals (Du\u0308rnegger and Haas (15)), but no absolute values have been found in the open literature. Therefore, the minimum film thickness measured in oil-lubricated seals is used. The contact zone geometry is defined by the parameters in Fig. 9b, where the film thickness h(x) is defined over the contact width b by the heights h1 and h2. Here three situations can be distinguished\u2014h1 < h2, h1 = h2, and h1 > h2\u2014which all may occur in practice depending on seal design and running conditions. Consequently, h2 is chosen as 2 \u00b5m such that the minimum film thickness is 1 \u00b5m when h1/h2 = 2. For the thin lubricant films and high shear rates in the contact zone, the rheology model has to be extrapolated to shear rates of \u03b3\u0307 = 1\u00b7107 s\u22121. This extrapolation has to be done with great care and will be reviewed in the discussion of the model results",
+            " Subsequently the system will heat up due to frictional heating in the contact. This frictional heating is not incorporated in the model, and steady-state situations with constant temperature have been considered only. Du\u0308rnegger and Haas (15) measured a temperature of 57\u25e6C on the shaft surface for a grease-lubricated seal at 10.5 m/s. This temperature was measured close to the contact, and temperatures in the contact itself are higher. Local temperatures of 70\u2013120\u25e6C are expected in the lip contact. The vicinity of the contact contains grease as well (see Fig. 9a). Shear rates are much lower here than in the contact zone, but due to the large width a + d, a lift force will be generated by the normal stress effect. The shear rate depends on the shaft speed u and the geometry that is defined by the air side and lubricant side angles \u03b1 and \u03b2. The contact width b, as defined in Fig. 9b, is set at zero here so that the influence of the contact zone and the vicinity of the contact can be studied separately. To obtain the total picture of the lift generated, both effects can be combined. The seal lip model includes the lip geometry h(x), which is defined by the parameters in Fig. 9 and is used to calculate the local shear rate under the lip as \u03b3\u0307 = u h(x) , [12] D ow nl oa de d by [ G eo rg ia T ec h L ib ra ry ] at 1 0: 05 1 2 N ov em be r 20 14 where u is the shaft surface speed. Equation [12] can be substituted into Eq. [4] and Eq. [7] to obtain the shear stress and normal stress as a function of the axial position x. The specific lift force Flift, which is the result of the normal stresses in the contact and vicinity of the lip, can be calculated by integrating the normal stress difference: Flift = \u222b x2 x1 N1(x)dx, [13] where x1 and x2 represent the zone of interest",
+            " For the model a contact width of a run-in seal of 200 \u00b5m is assumed. Three different contact geometries are studied: a flat contact zone with constant film thickness h1 = h2 = 2 \u00b5m and a contact zone with a minimum or maximum film thickness at c/b = 1/3 such that h1 = h2. In case of the flat contact zone, the lift is proportional to the width b. The geometry parameters for the seal have been chosen based on realistic values found in open literature. Values for the initial choice of the seal geometry, Fig. 9, can be found in Table 3. Figure 10 shows the results of the specific lift force Flift for different temperatures and contact zone geometries at 70\u25e6C. It can be seen here that small variations in contact geometry h1/h2 have a relatively small influence on the total lift force, and the effect of temperature appears to be more significant here. The model predicts a specific lift force of \u223c2 N/m for the flat contact at 70\u25e6C. This value can be D ow nl oa de d by [ G eo rg ia T ec h L ib ra ry ] at 1 0: 05 1 2 N ov em be r 20 14 TABLE 3\u2014INITIAL CHOICE OF SEAL GEOMETRY PARAMETERS Seal Geometry Parameters b 200 \u00b5m c/b 1/3 \u2014 h1 2 \u00b5m h2 1\u20133 \u00b5m \u03b1 20\u201340 \u25e6 \u03b2 30\u201360 \u25e6 Flip 6 N/m related to the seal\u2019s specific lip force, which can vary from 6 N/m for low-contact pressure-bearing seals to 60 N/m for shaft seals that have a garter spring",
+            " This lift is independent of the seal surface roughness, which is normally generated during runningin, and therefore the normal stress effect may already generate a lubricating film when the seal is new. In the vicinity of the contact the grease also contributes to the lift. Despite the relatively low shear rates here, the contribution of the normal stress effect is significant because of the relatively large widths a and d. The temperature in the vicinity of the contact is expected to be lower than in the contact zone, which increases the normal stress effect. Figure 9a defines the lip geom- Fig. 10\u2014Lip contact model for different temperatures and contact geometry h1 = 2 \u00b5m and b = 200 \u00b5m. etry where grease is present at both sides of the contact with a maximum height hg. Figure 11 shows the model results for a shaft speed u = 10 m/s and a minimum film height ho = 2 \u00b5m. The contact zone b is not included here such that b = 0 \u00b5m, and the calculated lift force is the result of the grease in the vicinity of the contact only. Figure 11 shows that the grease in the vicinity of the contact generates a lift of \u223c1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001170_1.2976450-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001170_1.2976450-Figure5-1.png",
+        "caption": "Fig. 5 Similar irregular hexagons \u201etop view\u2026",
+        "texts": [
+            " For the Gough\u2013Stewart platforms with a coplanar base only or with an irregular hexagonal base and platform, coefficients f1 , f2 , f5, and f11 will vanish and the obtained singularity equations will take the following form: f3x2z + f4x2 + f6xyz + f7xy + f8xz2 + f9xz + f10x + f12y2z + f13y2 + f14yz2 + f15yz + f16y + f17z 3 + f18z 2 + f19z + f20 = 0. 14 This form is also consistent with the one given in Ref. 5 . Furthermore, for the Gough\u2013Stewart platforms with similar irregular hexagons, as shown in Fig. 5, the total number of geometric parameters will reduce to 10 because t10 / t1= t11 / t2= t12 / t3 = t13 / t4= t14 / t5= t15 / t6= t16 / t7= t17 / t8= t18 / t9=rt. In this case, most coefficients except f10, f16, f19, and f20 will vanish, and the obtained singularity equations will take the following form: f10x + f16y + f19z + f20 = 0 15 3.3 SSM, TSSM, and Semiregular Hexagons. A SSM is shown in Fig. 6. A TSSM is shown in Fig. 7, and the Gough\u2013 Stewart platform with semiregular hexagons is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003636_s00542-012-1544-7-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003636_s00542-012-1544-7-Figure3-1.png",
+        "caption": "Fig. 3 Finite volume model and boundary conditions of the outlet field",
+        "texts": [
+            " And the oil leakage of operating FDBs is dominantly determined by the outlet flow near the air\u2013oil interface. So we modeled the outlet field which is composed of grooved thrust bearing, plain journal bearing, and air\u2013oil interface to investigate the motion of fluid lubricant and air\u2013oil interface due to seal design. The coupled effect between the outlet field and internal field was assumed to be very small in the steady state of this simulation. The finite volume model and boundary conditions are shown in Fig. 3. The finite volume model of the FBDs has 547,018 four-node tetrahedron cells. Because the spiral thrust bearing in this model has 20 grooves which repeat the geometrical pattern in every 18 , the periodic boundary condition was applied to both circumferential surfaces in order to reduce the computation time. Because the inlet of FDB Leakage Contamination Leaking oilFig. 1 Contamination problem of the HDD due to the leaking oil of FDBs the model is connected to the plain journal bearing of internal field and the pressure difference of the plain journal bearing is very small at steady-state, the mass flow due to pressure effect and centrifugal effect from the plain journal bearing to the inlet of the simulation model is very small",
+            " This simulation used the densities and viscosities of the fluid lubricant and air, surface tension coefficient of fluid lubricant with respect to air, and static contact angle at a temperature of 20 C. We compared the steady-state solutions of the Navier\u2013 Stokes periodic model with those of the Reynolds model (Jang et al. 2006) to verify the accuracy of the proposed model numerically. This simulation used the thrust grooved bearing area with small gap of the Navier\u2013Stokes periodic model as shown in Fig. 3 which satisfies the Reynolds assumption. The pressure boundary conditions at the inlet and outlet are typically used to solve the Reynolds model. The boundary condition of atmospheric pressure was applied to inlet and outlet of the Reynolds model and the Navier\u2013Stokes periodic model only for the purpose of numerical verification. The simulated pressure distribution and load capacity of the Navier\u2013Stokes periodic model match well with those of the Reynolds model as shown in Fig. 4. The axial load capacity of the Navier\u2013Stokes periodic model (7"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.24-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.24-1.png",
+        "caption": "Fig. 1.24. Influence of steering on the stability margin (system of Fig. 1.20 (bottom)).",
+        "texts": [
+            " An increase in 6 (but also an increase in spee~ V) reduces the stability margin until it is totally vanished as soon as the two singular points merge (also the corresponding points I and II on the handling curve of Fig. 1.17) and the domain breaks open. As a result, all trajectories starting above the lower separatrix tend to leave the area. This can only be stopped by either quickly reducing the steer angle or enlarging 6 to around 0.2rad or more. The latter situation appears to be stable again (focus) as has been stated before. For the understeered vehicle of Fig. 1.24 stability is practically always ensured. For a further appreciation of the phase diagram it is of interest to determine the new initial state (ro, Vo) after the action of a lateral impulse to the vehicle (cf. Fig. 1.25). For an impulse S acting at a distance x in front of the centre of gravity the increase in r and v becomes: S x S A r - , A v - (1.98) I m which results in the direction a A r x a b ,dv b k 2 (1.99) TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 51 The figure shows the change in state vector for different points of application and direction of the impulse S (k 2 = I/m- ab)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002774_s12206-011-1203-4-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002774_s12206-011-1203-4-Figure4-1.png",
+        "caption": "Fig. 4. Finite element model of bolt connection.",
+        "texts": [
+            " The amplitude refers to the difference between the maximum stress under the maximum working load and the maximum stress under the minimum load. The orthogonal design of the bolt is based on ANSYS. In order to increase the efficiency, a symmetrical treatment is adopted to establish the finite element model. In other words, only half of the model is computed and the symmetry constraints are applied in the symmetric plane. The loading surfaces are the contact surface of the bearing rings with bolt connecting and rolling bodies. The specific load conditions are shown in Fig. 4. According to the actual loading condition, the working load of the single bolt can be obtained by the formula below [9]: \u239f \u23a0 \u239e \u239c \u239d \u239b +\u22c5 \u22c5 = z F z M COS F axT A \u03c0 \u03c0\u03b2 sin R 21 (1) where \u03b2 refers to the included angle between the working load direction and the axial direction, z to the number of the bolts that connect the outer ring of the bearing, and R to the radius of the node in the rolling body. First, the finite element method is adopted to compute the stress of the bolt with different combinations of factors at different levels"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001383_sensor.2007.4300334-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001383_sensor.2007.4300334-Figure4-1.png",
+        "caption": "Fig. 4: Fixture and set-up used to test turbine devices.",
+        "texts": [
+            " This process was detailed in [7] and is a critical step in the fabrication process that is not accomplished using conventional wafer-to-wafer bond aligning means. After bonding, DRIE is used to etch the turbine blades on one side and the reference geometry on the opposite side of the rotor using the previously patterned silicon dioxide mask. These patterns are etched to a depth of 200 \u00b5m meeting the silicon race and thereby releasing the rotor (Fig. 3). A fixture formed using a Stratasys FDM Titan rapid prototyping tool (Fig. 4) enables testing without the need to integrate fluid connections onto the silicon device. The turbine side faces down in a pocket having a 125 \u00b5m clearance below the turbine blades permitting them to rotate without interference. Inlet and outlet nitrogen ports butt up against the inlet and outlet channels etched into the device during the final DRIE step. An o-ring is placed between the top of the device and the lid which is clamped down using screws. Slots are fabricated into the lid allowing a Philtec Silicon Dioxide Silicon (b) (c) AuSn (a) Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000284_acc.2006.1657378-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000284_acc.2006.1657378-Figure6-1.png",
+        "caption": "Fig. 6. Slave v and \u03b8 directions",
+        "texts": [],
+        "surrounding_texts": [
+            "To cope with the time-varying delay, two approaches are possible: a) the estimation of the plant state that was explained in section II. C and, b) the addition of a buffer d [3]. The buffer may be used to save the information that arrives from the opposite side of the teleoperation loop during a time that exceeds the maximum time delay. This information is then feed into the controllers at a constant rate to each controller. By using this method, the time delay may be kept constant at the expense of making it larger. For the teleoperation experiment, we implemented the buffer idea. However, our control law is capable of producing acceptable performance of the NCS even in the absence of the buffer, and without any other time-varying delay compensation scheme, as shown in Figures 7, and 8, which show the tracking performance of the remote slave robot. The force reflected to the master in this experiment is due to the dynamics of the slave robot. In other words, even though there is no force applied to the robot from an obstacle, gravity, friction and time delays force the robot to have a settling time different from zero. This in turn produces an error between the references sent to the slave and the actual state measurements, which forms the basis for the force reflection control laws designed in equations 5-8. The buffer size was chosen of 20, so this imply a constant time-delay in the loop of almost 20 times the average delay. This size was chosen assuming that neither the network nor the computer processing time will induced any longer delay and it worked reasonable for the experiment. Since the delay time in the loop was incremented by the buffer inclusion, the control gains were tuned again. In the case of the robot control law , the gains were as follows: Krv = 100.0, B\u03c6\u03b8 = 1\u00d7105 and K\u03c6\u03b8 = 2.5\u00d7106. For the haptic device control law, the gains were: Br = 0.5, Kr = 0.001, Krv = 100.0, B\u03c6 = 0.1, B\u03c6\u03b8 = 100.0, K\u03c6\u03b8 = 2500.0 and \u03bb = 0.04. The experimental obtained by using the buffer are shown in Figures 9 to 12. From these results we see that the tracking in velocity and angle experience a longer delay, caused by the buffer. However, the tracking in the angle is more accurate than when the buffer was not used."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001077_detc2009-87092-Figure17-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001077_detc2009-87092-Figure17-1.png",
+        "caption": "Figure 17. Experimental setup [5].",
+        "texts": [
+            " Bearing stiffness is slightly nonlinear, particularly under a small torque. Torsional rigidity of the cycloid drive is analyzed by measuring housing angular displacement under applied torque. Figure 16 shows a torque and angular displacement graph from torsional rigidity analysis is shown. Torsional rigidity decreases under torque that is much smaller than rated torque because of nonlinear Hertz contact and bearing rigidity. In order to verify the torsional rigidity analysis, an experimental setup was built as shown in Figure 17 [5]. While the input shaft is fixed with brake, a load bar attached to the output and 100 kgf preload was applied at the end of the load bar. As force at the end increased up to 200kgf upward, the force and the angle displacement were measured with load cell and output encoder (Heidenhain, RCN226), respectively. The experimental result is shown in Figure 18, which shows clearly the advantages of cycloid drive such as small backlash and lost motion. In addition, calculated torsional rigidity (102"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002411_tpwrd.2011.2176967-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002411_tpwrd.2011.2176967-Figure2-1.png",
+        "caption": "Fig. 2. Node i connected at segment k.",
+        "texts": [],
+        "surrounding_texts": [
+            "This is the force created by the friction between the air and the conductors. Applied in each node is the half sum of the friction forces of the joined segments connected to it (i.e., [10]) (11) where is the frictional force vector in segment at time and it depends on the relative velocity of wind on each segment (12) where frictional coefficient of the segment , at time ; diameter of segment ; air density; perpendicular to segment component of the wind velocity vector, at time ; perpendicular to segment component of velocity vector of segment , at time ; unit vector in the direction perpendicular to segment . Each segment is considered to have a velocity equal to the average of its extreme nodes\u2019 ( and ) velocities (13) The frictional coefficient is obtained from Fig. 3 by means of the Reynolds number (14) where is the air dynamics viscosity."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001236_j100686a017-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001236_j100686a017-Figure6-1.png",
+        "caption": "Figure 6. PMMA gels in benzene, plotted according to eq 12a. Freezing point depressions of solution-crosslinked",
+        "texts": [
+            " This is borne out by our experimental data for the W series, which show a straight line passing through the origin if AT is plotted against t#I/(l - 4). From the slope a value for u of 1.42 ergs/cm2 can be calculated, indicating that the interfacial energy (or interfacial area) may well be dependent on chain conformation. Similar plots for the X, Y, and 2 series also exhibit straight lines, but with progressively larger slopes and intercepts on the abscissa indicating that the simplifying assumptions underlying eq 12a become less justified when the amount of diluent during crosslinking is reduced (see Figure 6). A probable explanation may be provided by the effect of the increasing number of polymer-polymer segment contacts in these systems, which will affect the value of u and introduce an additional term on the right-hand side of eq 12a. The freezing point depressions of the gels according to eq 2 is shown in Figure 7 where AT is plotted against l/a. Here a is assumed to be equal to (?)\u2019\u201d and calculated from M , using (13) \u2019/a -2 \u2018/% (p)\u2019/z = 4 d Q o ) (To 1 a is the molecular expansion factor, which, according to Stockmayer and Fixman,20 equals a3 - 1 = 22 (14) where the perturbation parameter x , to a first approximation, is given by z = 2(a/,,)a\u20192~a-\u2019(ro2/M,) -a\u2019zA~Mc1\u20192 (15) where (2)\u2019\u201d is the root mean square end-to-end distance of the unperturbed chain of length M,"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002850_0022-4898(65)90022-4-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002850_0022-4898(65)90022-4-Figure3-1.png",
+        "caption": "FIG. 3.",
+        "texts": [
+            " The first was performed by means of bevameters (i.e. measuring devices to define the correlation between acting external forces and soil deformations caused by them). The second part dealt with measurement of drawbar-pull of a scale model pneumatic tyred vehicle. The third one contained traction measurement of a rigid wheel with lugs made in the two different scales. The measurement of soil parameters On the basis of the bevameter conception [2, 4, 9] two laboratory devices were made. Using the apparatus shown in Fig. 3, the pressure sinkage relationship is obtained by pressing rectangular footings into the soil. The experimental values of Schematic drawing of the bevameter, the test device for the determination of the load-sinkage relation of the soil. 34 A. SOLTYNSKI soil parameters K~ and K, are found using two footings of different widths b ~ = l l - 0 mm and b_,=15-5 ram, both plates being the length /=6 2 ram. From equation (9) K~= b~b2 ( a ~ - ~ ) b,. - bl K~= a,.b2-a,bl b~ - b~ where aa and a~ are the corresponding values of the vertical coordinates taken from Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002648_j.mechmachtheory.2011.11.014-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002648_j.mechmachtheory.2011.11.014-Figure5-1.png",
+        "caption": "Fig. 5. The hypothetical CRRC linkage for two adjacent polygonal links.",
+        "texts": [
+            " There are 24 ternary links for this linkage and indeed the supporting polyhedron is not amere snub disphenoid, but a double snub disphenoid. In a double polyhedron faces overlap pairwise. This is a standard trick to mobilize polyhedra with even valence known since [5]. When IRAs of neighboring faces intersect each other we can think of the motion of two neighboring rays (from rotation centers to concurrent corners of adjacent polygonal links) as the links of a CRRC linkage where the axes of C joints and also R joints intersect and adjacent C and R axis are parallel. In Fig. 5, O12 and O23 are the homothety centers of neighboring polygonal faces and E2 is the common corner of the corresponding polygonal links. E12 moves along the intersection of two cylinders along IRAs n12 and n23 with radii r12 and r23. In general two cylinders intersect along a 4th order spatial curve. Consider the projection 2 To avoid confusion we use corner for polygons and vertex for polyhedra. of the intersection curve on the plane of the IRAs (xy plane in figure). The components of any point (x, y) of this projection curve on n12 and n23 give s12 and s23 \u2013 the distances of polygonal links/faces to the IRA intersection O. For a dilative motion s12/s23 should be invariant at all times, which is possible only if (x, y) is on a line and we get a line only if path of E2 is planar. Two cylinders with intersecting axis of revolution intersect along a planar curve only if they have the same radii. So we conclude that the perpendicular rays meeting at an edge should have the same length, i.e. r12=r23 in Fig. 5. Consider an n-valent (n is even) vertex figure cut along edge VE1 and developed into a plane as in Fig. 6. If the perpendicular rays meeting at an edge have the same length, the plane angles \u03c112, \u03c123, \u2026, \u03c1n1 at vertex V can be dissected as \u03c112 \u00bc \u03b11 \u00fe \u03b12; \u2026;\u03c1n1 \u00bc \u03b1n \u00fe \u03b11: In matrix form these equations read 1 1 0 \u22ef 0 0 1 1 \u22ef 0 \u22ee \u22ee \u22ee \u22ef \u22ee 1 0 0 \u22ef 1 2 664 3 775 \u03b11 \u03b12 \u22ee \u03b1n 2 664 3 775 \u00bc \u03c112 \u03c123 \u22ee \u03c1n1 2 664 3 775 It can be shown by induction that this coefficient matrix has rank n\u22121 when the dimension n is even, so the plane angles \u03c112, \u03c123, \u2026, \u03c1n1 cannot be arbitrary, but one of them must be depending on the others",
+            " Next consider a contracted version of the base polyhedron by a ratio of cos\u03b8. As demonstrated in the previous section, the polygonal links constructed for the maximal configuration can be located into the new faces by rotating the links by \u03b8 about Pi. The sense of rotation of neighboring links is opposite to each other and the links are kept parallel to their original positions by means of the dap links. Meeting of the polygonal links is guaranteed by the symmetry condition, as we have explained above via Fig. 5. This relocation of the links is possible for any angle \u03b8, so the assembly is mobile with the motion parameter \u03b8. \u25a0 Also we have the following corollary: Corollary 2. If the homothety centers of any pair of neighboring faces are in symmetrical position with respect to the edge along the faces meet, the homothety centers on the faces around a vertex are equidistant to the vertex and the plane angles around a vertex can be dissected such that Eq. (1) is satisfied. Proof. Consider two homothety centers Oij and Ojk on neighboring faces in symmetrical position with respect to edge ej",
+            " It is here shown that the construction method given by R\u00f6schel [7] is the only way to obtain homothetic Jitterbug-like linkages with fixed IRAs. As the most symmetric cases, Platonic solids and the Archimedean duals, i.e. the Catalan solids have insphere. As an example, a tetrakis hexahedral linkage with the IRAs shown as guides is given in Fig. 12. In this linkage, the homothety centers on the faces are chosen as the points of tangencies with the insphere. Note that since tetrakis hexahedron is an isohedron any point on a face can be used as the homothety center by Theorem 3 (See Fig. 5 of [15] for a tetrakis hexahedral linkage for which the IRAs are not fixed throughout the motion). It is still possible to obtain homothetic linkages if there is no inscribed sphere, as our fruitful example of Fig. 4 demonstrates. Consider a 4-valent vertex figure, summation of opposite plane angles of which are equal (\u03c112+\u03c134=\u03c123+\u03c141) (Fig. 13). Choose a point O12 on face 12 and locate its mirror images O41 and O23 along edges e1 and e2, respectively. The mirror images of O23 and O41 along e3 and e4, necessarily coincide due to the condition \u03c112+\u03c134=\u03b11+\u03b12+\u03b13+\u03b14=\u03c123+\u03c141"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003888_s00021-012-0105-2-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003888_s00021-012-0105-2-Figure4-1.png",
+        "caption": "Fig. 4. Sketches of the general appearance of barrier solutions. The relative positions of IV and of V interchange according to whether or not V extends to \u03a01",
+        "texts": [
+            " These statements are not difficult to prove formally, the details amounting essentially to an exercise. Alternatively, the statements will be clear\u2014to some extent implicitly\u2014from the context of the material in the later sections. Explicit formal details can be found in [5]. A full characterization of behavior requires a finer delineation of cases. Let us fix attention on the plate \u03a02, and on an arbitrary angle \u03b32, in the range 0 \u2264 \u03b32 < \u03c0/2. We will have use for the five solution curves indicated by Roman numerals in Fig. 4, in which all illustrated curves are solutions of (2.4) meeting \u03a02 in angle \u03b32, as shown. We make the preliminary observation that if the starting point u2 on \u03a02 is high enough then in view of the geometrical interpretation of (2.4) relating curvature to height, the solution curve will bend sharply and become vertical prior to reaching \u03a01. By lowering u2 continuously we reach a (unique) level at which \u03a01 is just attained, at contact angle \u03b31 = 0. Continuing in this way, we see that \u03b31 then increases monotonically as u2 decreases until the contact angle \u03b31 = \u03c0 is attained on \u03a01",
+            " The apparently limited class of solutions envisioned by Laplace encompasses in fact all solutions expressible as graphs over a horizontal base domain that can appear; all solutions of the boundary problem (3.1) in two dimensions appear also as solutions of the corresponding one-dimensional problem (2.4), and we may safely limit ourselves to that simpler problem. In what follows we thus restrict discussion of solutions between the plates to solutions of (2.4b), with \u03c81 = \u03b31 \u2212 \u03c0/2,\u03c82 = \u03c0/2 \u2212 \u03b32 (3.2) The particular solutions noted in Fig. 4, corresponding to 0 \u2264 \u03b32 < \u03c0/2, are characterized by the properties: I. The upper barrier S0+, asymptotic to the x-axis at x = \u2212\u221e and meeting \u03a02 in angle \u03b32. II. The unique solution of (2.4), meeting \u03a01 on the x-axis, and meeting \u03a02 in angle \u03b32. III. The unique solution of (2.4), meeting \u03a02 in angle \u03b32 and crossing the x-axis at the midpoint between the two plates. This solution is symmetric relative to the midpoint and meets \u03a01 in the angle \u03b31 = \u03c0 \u2212 \u03b32. IV. The unique solution of (2.4), meeting \u03a02 in angle \u03b32 and \u03a01 in angle \u03b31 = \u03c0. V. The lower barrier S0\u2212, asymptotic to the x-axis at x = +\u221e and meeting \u03a02 in angle \u03b32 Figure 4a sketches the case in which 2a > x3 \u2212 x2, so that V does not extend to \u03a01. In the procedure of moving u2 downward when 0 \u2264 \u03b32 < \u03c0/2, the solution IV is then encountered prior to reaching V. If 2a < x3 \u2212 x2 then the order in which IV and V are encountered is reversed, as sketched in Fig. 4b. The barriers I, II, III however always appear in the order as written. The order in which these curves appear is important for the classification; we address this point in Sect. 5 below. We base our discussion on the result of [5] that the net force between the plates is attracting if and only if the trajectory S has a positive minimum or negative maximum u0, and that the horizontal force per unit length of the interface is then given by the explicit expression F = \u03c3\u03bau2 0. (4.1) We may assume we have a positive minimum, at (x0, u0)",
+            " We have given in (2.8) an explicit representation for that surface, and shown that S0+ 2 in conjunction with pressure forces exerts a horizontal force on \u03a02 exactly balancing the force due to any surface on the opposite side of \u03a02 and extending to the rest level u = 0 at x = +\u221e. Thus, such a configuration leads to no net force on \u03a02, regardless of \u03b32. We introduce correspondingly the reference surface V: S0\u2212 2 , meeting \u03a02 at the same angle and extending from a height y0\u2212 2 < 0 to vanishing height at x = +\u221e, see Fig. 4. We compare S0+ 2 with a surface S2: u2(x) meeting \u03a02 in the same angle \u03b32 but in a point y2 = y0+ 2 . As we have seen in the proof of Theorem 2.1, u2(x) cannot extend as a graph to an unbounded interval. If y2 > y0+ 2 then u2 (x) > y0+ 2 + \u03b4 (x2) > 0, and thus by (2.4) its curvature k > \u03b4(x2)/\u03ba > 0. Since u2(x) > u0+ 2 (x) > 0, this function will have a positive minimum u02, and since its curvature exceeds that of a circle of known radius, it must become vertical at a distance from the minimum point not exceeding that radius",
+            " In a repelling configuration, the total net horizontal force between the two plates is F0 = 2\u03c3 (1 \u2212 cos\u03c80) (5.5) We have thus reduced the problem to that of determining the angle \u03c80, in terms of the prescribed data and the distance 2a between the plates. We start by examining in general terms what happens, as we decrease the heights u2 from the height u0+ 2 of S0+ to the extent possible, while preserving the repelling property of the solution. Formally we can move the starting point downwards to the intersection height u0\u2212 2 of S0\u2212 with \u03a02, while keeping the intersection angle unchanged (see Fig. 4). In that way, when that limiting initial height is achieved, the solution will coincide with S0\u2212 and yield zero force, while all solutions with lower initial points on \u03a02 will produce attracting forces. There can be a difficulty with that procedure, as S0\u2212 may not extend to \u03a01, and thus an interval of the solutions u2(x) with initial points close enough to u0\u2212 2 will be physically unrealistic (see Fig. 4 and the discussion below). We address this issue by introducing the (unique) solution IV: w2(x; a) of (2.4) meeting \u03a02 in angle \u03b32 and \u03a01 in angle \u03c0. Then the range of admissible starting points on \u03a02 for repelling solutions can be taken to be max { w2 (x2; a) , u0\u2212 2 (x2) } < u2 (x2) < u0+ 2 (x2) (5.6) If u2 = u0+ 2 then by the general uniqueness theorem the entire trajectory of u(x) coincides with that of S0+, and the inclination \u03c802 = 0 is achieved at an idealized crossing point x02 = \u2212\u221e. As expected, (5",
+            " We observe further that during the procedure as described above, the contact angle \u03b31 increases monotonically through the range of achievable values, from \u03b30+ 1 to \u03b30\u2212 1 . This property permits us in much of what follows to replace the height parameter u2 by the contact angle \u03b31. We can make this procedure quantitative, so as to determine numerically which value on the left side of (5.6) is to be used. The locus S0+ becomes vertical at an abscissa x3 > x2, beyond which it cannot be continued and at which u3 = \u221a 2/\u03ba. Denoting by x4 the abscissa of the vertical point of S0\u2212, we see from symmetry properties of the equation that x2 \u2212 x4 = x3 \u2212 x2. (see Fig. 4). Thus the criterion that u0\u2212 (x2)should be replaced by w(x2) in (5.6) is that x3 \u2212 x2 < 2a. We proceed to estimate x3 \u2212 x2. Using (2.8) we are led to x3 \u2212 x2 = 1\u221a 2\u03ba \u03c0/2\u222b \u03c82 cos \u03c8\u221a 1 \u2212 cos \u03c8 d\u03c8 = 1\u221a 2\u03ba \u03c0/2\u222b \u03c0 2 \u2212\u03b32 \u221a 1 + cos \u03c8 cot \u03c8d\u03c8. (5.7) We obtain Theorem 5.2. The left side of (5.6) is to be chosen as u0\u2212 (x2) if and only if 2a \u2264 2a\u2217 .= 1\u221a 2\u03ba \u03c0/2\u222b \u03c0 2 \u2212\u03b32 \u221a 1 + cos \u03c8 cot \u03c8d\u03c8. (5.8) The integral in (5.8) is easily computed numerically. As a guide in individual cases, we note that \u2212 ln cos \u03b32 < \u03c0/2\u222b \u03c0 2 \u2212\u03b32 \u221a 1 + cos \u03c8 cot \u03c8d\u03c8 < \u2212 \u221a 1 + sin \u03b32 ln cos \u03b32 (5",
+            " It is the height on \u03a02 of the uniquely determined solution of (2.4) in the interval x1 < x < x2, which meets \u03a01 in angle \u03b31 = \u03c0 and \u03a02 in the prescribed angle \u03b32. We have need for this solution in the present context exactly in the situation for which (5.8) fails. We will have to distinguish cases, according to whether or not this locus crosses the x-axis between the plates \u03a01 and \u03a02. We start by determining the plate separation 2a0 dividing these cases, that is the separation for which w2 = 0 (see Fig. 4). Then the crossing angle with the x-axis becomes \u03c80 = (\u03c0/2 \u2212 \u03b32); integrating (5.1) from the contact point with \u03a01 (with \u03b31 = \u03c0) to the crossing point on the x-axis we find sin \u03b32 = 1 2 \u03baw2 1. (5.10) Using (2.7) adapted to this configuration we determine 2a0 = 1\u221a 2\u03ba \u03c0/2\u222b \u03c0 2 \u2212\u03b32 cos \u03c8\u221a sin \u03b32 \u2212 cos \u03c8 d\u03c8 (5.11) Since the separation 2a decreases with decreasing height w2, we see that w2(x; a) cuts the x-axis between the plates if and only if a > a0. We thus have reduced the problem to two explicit cases, according to the prescribed separation 2a and the explicitly given value 2a0",
+            " As indicated above, we are free to adopt \u03b31 \u2265 \u03b30+ 1 as parameter. In accordance with the Cases 1 and 2 indicated above, we distinguish configurations in which x0 < x1, x1 < x0 < x2, and x2 < x0. Within the family considered of solutions meeting \u03a02 in angle \u03b32, the cases are demarcated by the particular solutions joining the plates and meeting them on the x-axis. The former meets \u03a01 in an angle \u03b3\u2212 1 = \u03c801 + (\u03c0/2); the latter meets \u03a02 in the angle \u03b32 = (\u03c0/2) \u2212 \u03c802, and \u03a01 in the angle \u03b3+ 1 > \u03c0 \u2212 \u03b32 = \u03c802 + (\u03c0/2) > \u03c801 + (\u03c0/2) = \u03b3\u2212 1 (see Fig. 4). Adapting (2.7) to these cases, we find the relations 2a = 1\u221a 2\u03ba \u03c0 2 \u2212\u03b32\u222b \u03b3\u2212 1 \u2212 \u03c0 2 cos \u03c8\u221a sin \u03b3\u2212 1 \u2212 cos \u03c8 d\u03c8 (5.16) and 2a = 1\u221a 2\u03ba \u03b3+ 1 \u2212 \u03c0 2\u222b \u03c0 2 \u2212\u03b32 cos \u03c8\u221a sin \u03b32 \u2212 cos \u03c8 d\u03c8 (5.17) which determine the angle \u03b3\u00b1 1 in the respective instances. We are led to: Theorem 5.4. For \u03b32 fixed as above, suppose that \u03b31 is in the range Ia2 of repelling solutions. Let \u03b3\u2212 1 , \u03b3+ 1 be the respective solutions of (5.16), (5.17). If \u03b30+ 1 < \u03b31 < \u03b3\u2212 1 then the solution crosses the x-axis in the range \u2212\u221e < x < x1, at an angle \u03c80 such that 0 < \u03c80 < \u03c801",
+            " For each given (\u03b31, \u03b32) with \u03b32 \u2208 [0, \u03c0/2) and separation 2a, we subdivide Ia2 into three subsets: Ia2 = Il a2 \u222a Ir a2 \u222a Is a2 according to choice of \u03b31: Il a2 = those trajectories S2 \u2208 Ia2, for which \u03b31 + \u03b32 < \u03c0, and Ir a2 = those trajectories S2 \u2208 Ia2, for which \u03b31 + \u03b32 > \u03c0. Is a2 = the anti-symmetric solution with data (\u03c0 \u2212 \u03b32, \u03b32), for which \u03b31 + \u03b32 = \u03c0. It is clear that for each \u03b32 in the half-closed interval [0, \u03c0/2), each of the sets just introduced will be non-null. In this context we note that the traverse S2 that crosses the x-axis at the midpoint (x1 +x2)/2 between the plates is the single element of Is a2 \u2261 III, see Fig. 4 and Sect. 3. We consider the individual cases that can occur, and arrive at Theorem 5.4. If S2 \u2208 Il a2 \u222a Ir a2 then there exists a unique a\u2217 \u2208 (0, a) such that when the separation is reduced to 2a\u2217 the net horizontal force acting on the plates vanishes. For all smaller values of a\u2217 the plates attract each other, according to the laws developed in Sect. 4; notably the attracting force becomes unboundedly large as O ( 1/a2 \u2217 ) . For the unique S2 \u2261 III \u2208 Is a2 the force remains repelling for all plate separations, and is subject to the fixed upper bound 2\u03c3 in magnitude that is evident from (5.5). (In the following section we will establish more precise bounds on forces arising in repelling configurations.) Proof of Theorem 5.4. We introduce as reference surface the (unique) antisymmetric solution III: S0 2 : u0 2 (x; a) assuming the data \u03b30 2 = \u03b32 and \u03b30 1 = \u03c0 \u2212 \u03b32 on the respective plates, and crossing the x-axis at the midpoint 1 2 (x1 + x2) between the plates, see Fig. 4. Keeping \u03b31 and \u03b32 fixed and letting a \u2192 0 yields a family of solutions, all with the same fixed data on the plates, all crossing the x-axis at the midpoint, and tending asymptotically to a linear segment crossing the axis at the angle \u03c80 = (\u03c0/2) \u2212 \u03b32. Since all these solutions cross the axis, they yield repelling forces for all values of a, tending to the limiting force F02 = 2\u03c3(1 \u2212 sin\u03b32). Lemma 5.5. Let u(x) and v(x) be two solutions of (2.4) on a common interval I. Suppose that u\u2032 (x0) = v\u2032 (x0) at a point x0 \u2208 I",
+            " We have also introduced the symmetric barrier III: S0 2 : u0 2 (x; a), which serves to separate those repelling solutions that change to attracting solutions with decreasing a via S0+ with fluid rising between the plates, from those that pass through S0\u2212, and depress the fluid between the plates. We now introduce additionally the barrier solutions: 1. II: S1 2 : u1 2 (x; a)is the uniquely determined solution in Il a2, which passes through the point (x1,0) at the intersection of \u03a01 with the undisturbed fluid surface, and 2. II2: S2 2 : u2 2 (x; a) is the uniquely determined solution in Ir a2, which passes through the point (x2,0) at the intersection of \u03a02 with the undisturbed fluid surface. The former of these is indicated in Fig. 4, together with the four special solutions already introduced. The solution trajectories are mutually nonintersecting; the trajectories meeting \u03a02 in angle \u03b32 and extending to \u03a01 simply cover a region between the plates bounded between the curves that meet \u03a01 in the angle 0 or \u03c0. We label the barrier curves I to V, as indicated in the figure. All repelling solutions lie below I and above IV and V, throughout the intervals for which comparison is possible. We characterize these curves individually for later reference: I"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000469_14644193jmbd78-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000469_14644193jmbd78-Figure9-1.png",
+        "caption": "Fig. 9 Piston tilting",
+        "texts": [
+            " Thus, these components have Table 3 Internal and external forces applied No. Type Position Magnitude 1 Journal bearing reaction Left main journal bearing Refer to equation (11) 2 Journal bearing reaction Right main journal bearing Refer to equation (11) 3 Combustion gas force Perpendicular to piston top surface Figure 3 represents one complete cycle 4 Load torque Crankshaft end 30 Nm 5 Friction torque Journal bearings Refer to equation (13) 6 Friction at piston Piston top corners (1 and 2) as shown in Fig. 9 Refer to equation (16) 7 Normal forces at piston edge Piston top and bottom corners (1 to 4) as shown in Fig. 9 Refer to equation (15) been introduced into the model as flexible bodies, as their structural modes are more likely to be excited. The most important assumption behind this procedure is the consideration of small, linear body deformations relative to a local frame of reference, while this local frame of reference undergoes large, non-linear motions with respect to a fixed global frame of reference. The discretization of a component into a finite element model represents the infinite number of DOF with a finite, but very large number of DOF",
+            " Surface traction force (or the friction force) acting on the piston along the piston skirt opposing the piston translational motion is calculated, considering the viscous drag acting on the contact conjunction. Viscous drag force is given by Gohar [31] as \u03c4 = \u00b1 h 2 dP dx\ufe38 \ufe37\ufe37 \ufe38 a + \u03b70 U h\ufe38 \ufe37\ufe37 \ufe38 b (16) In hydrodynamic lubrication, the contribution from the pressure gradient on the viscous drag is very small as shown Gohar [31], so the component due to pressure gradient is neglected. The initial clearance h0 can be obtained as follows (for the contact point 1 in Fig. 9) h0 = xL \u2212 a sin \u03d1 \u2212 rtt cos \u03d1 (17) For simplicity, it is considered that only four corners of the piston touch the cylinder wall when the piston is in operation (due to its secondary motions, lateral and tilting) and these four positions are shown numbered in Figs 8 and 9 (points 1 to 4). This approach follows that of Haddad and Howard [35], who used the four corner model with spring-damper elements, which are clearly not representative of tribological conditions. Force component due to pressure variation is orthogonal to the piston surface at these four corners of the piston as shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001958_icinfa.2010.5512015-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001958_icinfa.2010.5512015-Figure6-1.png",
+        "caption": "Fig. 6 The least square line fitting algorithm sketch map.",
+        "texts": [
+            " Experiments prove that it will enhance the precision and reduce error if we choose Dthr as 15 percent of the length of line from the start point to the end point. 4) Line Segment Fitting: According to the above processing, we get several point sets Li ={P1, P2, \u2026.., Pn}, but these points have noise and have not strict linear relation. So we need a line segment to mirror the relationship between these points fully. This paper employs the least square line fitting algorithm proposed by Deriche to complete the line segment character parameter. In Figure 6, \u03c1 expresses the distance between the origin of coordinate and line L. \u03b1 is the angle included between the perpendicular from origin to line L and the X-axis. So point P(x, y) on the line L meets the following equation. (8) If id is the distance between point ( , )i i iP x y and the line segment L, the goal of least square line fitting algorithm is to find optimal ( , )\u03c1 \u03b1 , which causes to minimize the sum of the squares of the distances of the points from the line. cos( )i i id r \u03b1 \u03b8 \u03c1= \u2212 \u2212 (9) 2 2 ( , ) ( , )1 1 arg min arg min ( cos( ) ) (10) N N i i i i d r \u03c1 \u03b1 \u03c1 \u03b1 \u03b1 \u03b8 \u03c1 = = = \u2212 \u2212 The method of finding optimal ( , )\u03c1 \u03b1 is: 1 arctan 2 2 b a c \u03c0\u03b1 \u2217 = \u2212 \u2212 (11) cos sinx y\u03c1 \u03b1 \u03b1 \u2212 \u2212 \u2217 \u2217 \u2217= + (12) Including: 1 1 N i i x x N \u2212 = = (13) 1 1 N i i y y N \u2212 = = (14) cos sinx y\u03b1 \u03b1 \u03c1+ = 2 1 ( ) N i i a x x \u2212 = = \u2212 (15) 1 2 ( )( ) N i i i b x x y y \u2212 \u2212 = = \u2212 \u2212 (16) 2 1 ( ) N i i c y y \u2212 = = \u2212 (17) We obtain the optimal value ( , )\u03c1 \u03b1 of the line via least square line fitting algorithm"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002204_b978-0-08-003491-1.50027-7-Figure8-1.png",
+        "caption": "FIG. 8. Hydraulically driven spindle with hydrostatic bearings.",
+        "texts": [
+            " 7) is widely used for an extensive range of applications in precision form grinding, the machine being designed to grind to close limits complex contours such as occur on flat and circular form tools, crushing rolls, male and female profile templates and gauges, and press tool punch and die segments. Grinding Machine Spindles 461 With the aim of increasing the productivity of this machine while maintaining the high standards of accuracy and versatility new ideas have been examined. A development currently being evaluated is the hydraulically driven spindle with hydrostatic bearings shown in Fig. 8. The system it is proposed to replace consists of a conventional type of spindle supported on preloaded angular contact ball bearings. The drive is transmitted from an electric motor by belt and pulleys as shown in the photograph Fig. 9. In operation the spindle unit reciprocates on the vertical wheel slide, changes in the pulley position being compensated for by a spring loaded jockey pulley. This may be compared with the proposed new design of spindle in which the motor is incorporated as an extremely compact integral unit"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure10.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure10.4-1.png",
+        "caption": "Fig. 10.4-1 Engine mounts and coordinate systems.",
+        "texts": [
+            "3-10 Superposition and juxtaposition of confi guration graphs and direction cosine matrices. VibrationAnalysis_txt.indb 372 11/24/10 11:50:53 AM Engine Mounting Systems | Chapter 10 373 to the engine side so that the engine and mounts can be treated outside the context of a full vehicle; second, since mount optimization is usually needed in the early stages of the vehicle design process, it is desirable to optimize the mounting system before good defi nition of the remainder of the vehicle can be obtained. az of the elastic center of the mount (in Fig. 10.4-1 only one engine mount is shown). Th e translational displacements of the mass center of the engine are xc, yc, zc in the X, Y, Z directions, respectively. Th e rotational displacements of the engine are characterized by the angles , , of the engine body axes about the X, Y, Z axes, respectively. Only small translations and rotations are considered. Hence, the rotations are commutative (i.e., the resulting position is independent of the order of the component rotations) and the angles of rotation about the engine body axes are equal to those about the global X, Y, Z axes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003880_s00521-012-1326-2-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003880_s00521-012-1326-2-Figure6-1.png",
+        "caption": "Fig. 6 Two inverted pendulums on carts [8]",
+        "texts": [
+            " Example 2 Two inverted pendulums on carts Observing the design procedure of our approach, we can find that the control design is valid for nonlinear affine uncertain systems. Thus, to illustrate the effectiveness of our proposed approach, the proposed adaptive control scheme is applied for nonlinear affine system\u2014two inverted pendulums on carts. The comparison results with other method are also introduced to show the performance. Herein, we consider two inverted pendulums connected by a moving spring mounted on two carts as follows (show in Fig. 6). _x1 \u00bc x2 _x2 \u00bc g cl x11 g cl x12 2 64 3 75\u00fe 1 cml2 1 0 0 1 u\u00fe N21 N22 \" # \u00fe k\u00f0a\u00f0t\u00de cl\u00de cml2 \u00f0 a\u00f0t\u00dex11 \u00fe a\u00f0t\u00dex12 y1 \u00fe y2\u00de k\u00f0a\u00f0t\u00de cl\u00de cml2 \u00f0 a\u00f0t\u00dex12 \u00fe a\u00f0t\u00dex11 y1 \u00fe y2\u00de 2 664 3 775 \u00f039\u00de where x1\u00bc \u00bdx11;x12 T \u00bc \u00bdh1;h2 T ; x2\u00bc \u00bdx21;x22 T \u00bc \u00bd _h1; _h2 T , u\u00bc \u00bdu1;u2 T . The parameter values with appropriate units are given by g = l = 1, k = 1, M = 50, m = 50, y1 = sinx1t, y2 = sinx2t ? L, a(t) = sinx1t, L = 2. System unknown nonlinearity [N21 N22]T are N21\u00bc m M x2 21 sinx11; N22\u00bc m M x2 22 sinx12. External disturbance d is d \u00bc k\u00f0a\u00f0t\u00de cl\u00de cml2 \u00f0 a\u00f0t\u00dex11 \u00fe a\u00f0t\u00dex12 y1 \u00fe y2\u00de k\u00f0a\u00f0t\u00de cl\u00de cml2 \u00f0 a\u00f0t\u00dex12 \u00fe a\u00f0t\u00dex11 y1 \u00fe y2\u00de 2 664 3 775: \u00f040\u00de The initial conditions are \u00bdx11;x12;x21;x22 \u00bc \u00bdp6 ; p 6 ;0;0 : The following controller parameters of Theorem 1 are chosen k1 = k2 = 10, r1 = 200, r2 = r3 = r4 = 0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002341_j.optlastec.2011.04.010-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002341_j.optlastec.2011.04.010-Figure2-1.png",
+        "caption": "Fig. 2. Calculated results of the water film and its schematic simplicity.",
+        "texts": [
+            " To investigate these processes, three numerical models and the thermo-mechanical analyses are needed. The processes of LSM in air, under water film and under water are indicated in Fig. 1. Fig. 1(a) shows the general LSM in dry air as well as the substrate geometry and results measuring positions used in all cases. For water-assisted LSM in Fig. 1(b), water film was generated by a jetting nozzle with a radius of 1.5 mm, flow angle of 451 and speed of 0.3 m/s, which moved following the laser simultaneously. The distance from jetting incidence point to laser spot was set to 5 mm. As shown in Fig. 2, the water flowage area was calculated using the hydrodynamic code Fluent. In order to be induced into the calculation conveniently, the boundaries of this area were simplified and depicted by three equations. Workpiece was fixed into a tub of water for under water LSM as in Fig. 1(c), such that the steel sample was kept 2 mm below water surface. For assurance of the quality in under water laser processing, the effect of water on the metallurgical behavior should be considered [20]. In order to exclude the water around the remelting zone (RMZ), a local dry cavity on the workpiece was created approximately 5 mm in diameter by an air stream coaxial with laser beam [15]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure10.7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure10.7-1.png",
+        "caption": "Fig. 10.7-1 Torque roll axis.",
+        "texts": [
+            " Th erefore, if an engine is suspended by very low rate motor mounts, and if it is then subjected to an oscillating crankshaft torque, the engine will oscillate about its TRA. In the following paragraphs, we present a procedure for fi nding the direction of the TRA. Let X, Y, Z be an engine coordinate system with origin at the engine mass center, and with the X-axis parallel to the crankshaft axis. Let the Y-axis then be directed to its right when looking along the positive X-axis. Th e positive Z-axis is then determined by the right-hand rule. See Fig. 10.7-1. Let N1, N2, N3 be mutually perpendicular unit vectors parallel to the X, Y, and Z axes, and let n1, n2, n3 be mutually perpendicular unit vectors parallel to the engine\u2019s principal axes of inertia. Th e relation between these two axis systems can be expressed by an orthogonal transformation matrix of direction cosines as n n n 1 2 3 1 1 1 2 2 \u23a1 \u23a3 \u23a2 \u23a2 \u23a2 \u23a4 \u23a6 \u23a5 \u23a5 \u23a5 = cos cos cos cos cos c os cos cos cos 2 3 3 3 1 2 3 \u23a1 \u23a3 \u23a2 \u23a2 \u23a2 \u23a4 \u23a6 \u23a5 \u23a5 \u23a5 \u23a1 \u23a3 \u23a2 \u23a2 \u23a2 \u23a4 \u23a6 N N N \u23a5 \u23a5 \u23a5 , (10.7-1) where i, i , and i (i =1 2 3, , ) are orientation angles"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003879_20131120-3-fr-4045.00005-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003879_20131120-3-fr-4045.00005-Figure1-1.png",
+        "caption": "Fig. 1. Custom-made autopilot board",
+        "texts": [
+            " In Section 2 the autopilot and control architecture is described. A description of the small UAV is in Section 3. The guidance algorithm is presented in Section 4. The validation of this algorithm is performed via HIL simulations (see Section 5 and 6). The conclusions are summarized in Section 7. Usually commercial autopilots are not reconfigurable, this means that variations of on board software are not allowed. For this reason a custom-made autopilot, designed and produced by the researchers of Politecnico di Torino [4] (see Figure 1), is considered in this paper. Its main characteristics comprehend an open architecture, the possibility to be reprogrammed in flight and real time telemetry. Sensors include GPS, barometric sensor, differential pressure sensor and three-axis gyros and accelerometers. The CPU is the AtMega 1280 model with 128Kb flash memory and 8Kb of RAM with a CPU clock of 16 MHz. The software part consists of a multi-skilled algorithm in C language, designed for the management of: * flight data acquisition, * navigation algorithm, based on a Kalman filter for sensor acquisition and data fusion, * guidance algorithm, described in this paper and im- plemented to guide the UAV through waypoints with a feasible trajectory, * controller algorithm, for the inner and outer loop variables"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002616_demped.2011.6063683-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002616_demped.2011.6063683-Figure9-1.png",
+        "caption": "Fig. 9. Assembly of static eccentricity",
+        "texts": [
+            " For example for the first rotor slot harmonic does not appear because its order does not belong to , , on the contrary of the second rotor slot harmonic which appears at the frequency for Figure 8 shows the spectral content of the stator current of the machine with the presence of static eccentricity. The frequencies orders in this case are given by: , (32) For unlike the case where the machine is healthy, both RSH are present on the spectrum of the stator current. For the experimental part, as shown Fig. 9 a new housing with an eccentric additional ring is used to simulate the static eccentricity. The experimental given results in Fig. 10 confirm the simulation ones. IV. CONCLUSION A transient model of squirrel cage induction has been presented. This model takes into account the axial fault, and based on an analytical development of the different inductances formulas as Fourier series without any kind of reference frame transformation. On the other hand, the objective of the different mathematical developments for the model is to allow in plus the simulation results an analytical explanation of the causes and effects of different diagnosis signatures"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001954_13506501jet521-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001954_13506501jet521-Figure1-1.png",
+        "caption": "Fig. 1 Half of a plain journal bearing",
+        "texts": [
+            " A total of 12 identical plain journal bearings were used to investigate the influence of sea water on bearing performance. Each bearing had a different but fixed working condition. The amount of sea water in terms of mass changed from 0, 1, and 2 per cent to 3 per cent, and the speed of rotation varied from 500, 1000, and 2000 to 2750 r/min. The following performance parameters were measured: oil film temperature, journal loci, emulsion viscosity, oil film friction, and power consumption. A typical test bearing is shown in Fig. 1 with the following parameters given in Table 1. The inner wall of the Proc. IMechE Vol. 223 Part J: J. Engineering Tribology JET521 \u00a9 IMechE 2009 at PENNSYLVANIA STATE UNIV on May 23, 2015pij.sagepub.comDownloaded from bearing was built from tin-based Babbitt alloy, which is soft, wear resistant and heat conducting, and the bulk material was 40Cr. The composition of the tin-based Babbitt alloy was Sn 85.13 per cent, Pb 0.024 per cent, Cu 5.84 per cent, and Sb 10.61 per cent. The test set-up is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002504_s10015-011-0948-2-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002504_s10015-011-0948-2-Figure8-1.png",
+        "caption": "Fig. 8. Operation for translational motion",
+        "texts": [
+            " From this fi gure, we can see that the robot base can follow the posture of the operating device. In particular, it can be seen that the robot base can retain its position and posture in spite of the movement of the manipulator shown in Fig. 6b. Figure 7 shows the experimental results in the case of translational motion. The pictures in Fig. 7a and b were taken from the front of the water tank, and the picture in Fig. 7c was taken from the side of the water tank. In this operation, the robot base can be operated by using the three potentiometers in the base operating device shown in Fig. 8. We have proposed a master\u2013slave control system for a UVMS. Our proposed master controller of the vehicle can be manipulated with only one hand. The effectiveness of the proposed master\u2013slave control system was demonstrated by using a fl oating underwater robot with a 2-link manipulator. 1. Maheshi H, Yuh J, Lakshmi R (1991) A coordinated control of an underwater vehicle and robotic manipulator. J Robotic Syst 8:339\u2013370 2. McLain TW, Rock SM, Lee MJ (1996) Experiments in the coordinated control of an underwater arm/vehicle system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000224_jbm.a.30807-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000224_jbm.a.30807-Figure2-1.png",
+        "caption": "Figure 2. Representation of the J contour integral for estimating fracture toughness.",
+        "texts": [
+            " Methods of analysis Although there are several techniques for calculating a fracture parameter, the J contour integral is considered to be the most suitable because this method can be employed for specimens possessing arbitrary geometry.14 It can also be applied both to linear and nonlinear problems, and its path independence (in an elastic material) permits evaluation at a remote contour, where numerical accuracy is greater.15 Journal of Biomedical Materials Research Part A DOI 10.1002/jbm.a Consider an arbitrary counter-clockwise path ( ) around the tip of a crack, as illustrated in Figure 2. The J contour integral can be calculated from the expression J T Wdy T u x ds (1) where W is the strain energy density, T is the component of the traction vector, u is the displacement vector component, and ds is an incremental length the contour . The strain energy is defined as follows: W 0 \u03b5 d\u03b5 (2) where and are the stress and strain tensors, respectively. The traction is a stress vector normal to the contour and its components are given by T ijnj (3) where nj is the component of the unit vector normal to , xx is a normal stress in x direction, yy is a normal stress in y direction, zz is a normal stress in z direction, xy and yx are shear stresses in xy plane, xz and zx are shear stresses in xz plane, and yz and zy are shear stresses in yz plane"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002565_gt2011-46492-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002565_gt2011-46492-Figure1-1.png",
+        "caption": "Figure 1. Cross section of bristle pack: hexagonal close packed bristle",
+        "texts": [
+            " Computational Fluid Dynamics modelling of the bristle pack, drawing from previous research [9 \u2013 14], forms the basis of the parametric study, and the numerical outputs are processed to obtain key performance measures for comparison between the geometric configurations. Inferences are thus drawn from these data on the significant parameters in brush seal bristle pack design. A family of idealised brush seal domains was created using the commercial meshing package GAMBIT (Version 2.3.16). These comprised a single circumferential bristle row with constant interbristle spacing. The bristles were arranged into a hexagonal close-packed arrangement, as depicted in the cross-section in Fig. 1, which represents the physical arrangement of bristles in a real brush seal prior to pressurization. In a real brush seal, a differential pressure will cause deflection (and tighter packing) of the bristle elements, resulting in a slightly disconfigured bristle arrangement. Even though these deflections change the interstitial gap sizes along the bristle length, and lead to inter-bristle and bristle-backing ring contact, an idealised pack configuration gives a reasonable indication of the expected flow and temperature fields within the pack"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001558_lars.2008.26-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001558_lars.2008.26-Figure3-1.png",
+        "caption": "Fig. 3. Roll moments from flapping",
+        "texts": [
+            " (13) The stabilizer bar affects the vehicle dynamics exclusively by augmenting the cyclic pitch command to the main rotor blade. The lateral and longitudinal commands are rewritten as: U\u0304lat =Ulat +Kdd and U\u0304lon =Ulon +Kcd, (14) where Kc and Kd represent the gearing of stabilizer bar mixer. Defining the stabilizer bars coupling derivatives Bd = BlatKd and Ac = AlonKc (15) and using U\u0304lat and U\u0304lon instead of Ulat and Ulon in Eqs. (10) and (11), the main rotor flapping state equations become a\u0307 = 1 \u03c4 f [\u2212a\u2212 \u03c4 f q+Abb+Acc+AlonUlon +AlatUlat ], (16) b\u0307 = 1 \u03c4 f [\u2212b\u2212 \u03c4 f p+Bab+Bdd+BlonUlon +BlatUlat ]. (17) Fig. 3 depicts the relation between the rotor tip-path-plane lateral angle b, the thrust vector T and the moments produced by the main rotor. A similar analysis can be performed to the longitudinal angle. It\u2019s assumed that the thrust vector remains perpendicular to the rotor tip-path-plane in low speed flight. The components of the thrust vector are: Tx =\u2212|T|sin(a)cos(b) (18) Ty = |T|sin(b)cos(a) (19) Tz =\u2212|T|cos(a)cos(b) (20) These components act in the helicopter\u2019s fuselage with a moment arm h which is the distance from the hub to the center of gravity cg resulting in the moments above: MxT = hTy = h|T|sin(b)cos(a) (21) MyT =\u2212hTx = h|T|sin(a)cos(b)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000870_1.2747645-Figure15-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000870_1.2747645-Figure15-1.png",
+        "caption": "Fig. 15 Circular orbit of the shaft center at the slip rings",
+        "texts": [
+            " The real part of the thermal eigenvalue T indicates whether the spiral increases or decreases in magnitude: 0 Increasing magnitude unstable behavior A12a 0 Decreasing magnitude stable behavior A12b Transactions of the ASME 15 Terms of Use: http://asme.org/terms A E A t e F h J Downloaded Fr ppendix B: Estimation of the Ratio Between Added to liminated Heat at the Slip Rings ssumptions for the Estimation The analytical estimation is done with the following assump- ions: 1. The slip ring orbit is a synchronous circle Fig. 15 . The orbit at the slip rings is circular. 2. The circumferential distribution of the added heat due to the vibration is a sinus shape Fig. 16 . The maximum is at the high spot of the shaft. 3. For the eliminated heat distribution, the same applies as for the added heat. 4. The axial distribution of the temperature in the slip ring is constant. 5. The slip ring is adiabatic. 6. The complete sinusoidal shaped friction power due to the vibration enters the ring. 7. The thermal bow of the slip ring is fully transmitted to the shaft"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003779_s11071-012-0647-0-Figure17-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003779_s11071-012-0647-0-Figure17-1.png",
+        "caption": "Fig. 17 Comparison of convergence regions for the two controllers. Fuzzy-Pad\u00e9 convergence region is %348.74 larger than that of the fuzzy controller proposed by Yi et al. [34]",
+        "texts": [
+            "3 Results and comparisons To verify the effectiveness of the proposed fuzzy-Pad\u00e9 controller, several simulations are done. At first, Figs. 15 and 16 show fuzzy-Pad\u00e9 results compared to the fuzzy controller proposed by Yi et al. [53]. In Table 10, the two controllers are compared in terms of CPU-time and energy consumption of the system. Data in Table 10 belong to the simulations in Figs. 15 and 16. Table 10 shows that the fuzzy-Pad]\u00e9 controller is superior to the fuzzy controller designed by Yi et al. [53]. Figure 17 shows the convergence regions of the initial angle of the pendulums for both fuzzy and fuzzyPad\u00e9 controllers. It is shown that the convergence re- gion of the fuzzy-Pad\u00e9 controller is much larger than that of the fuzzy controller, %348.74 larger. Figure 18 shows the fuzzy-Pad\u00e9 simulation results with different initial angles, where the fuzzy controller solutions diverge in these cases. In Fig. 18, the left and the right axes separately represent the angle of the two pendulums and the position of the cart"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001793_j.sna.2010.11.014-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001793_j.sna.2010.11.014-Figure3-1.png",
+        "caption": "Fig. 3. (a) Fabrication sequence for modified glass substrate based on polymeric barrier layer and (b) setup for measuring transport velocity of droplet.",
+        "texts": [
+            " The four hydrophobic barrier layers, AF 601, C4F8, PDMS, and ITO, were patterned in the shape of a square y wet etching or a lift-off process. In order to pattern the ITO barier layer, wet etching was conducted using an ITO etchant, LCE-12 Cyantek). To pattern the three other polymeric barrier layers, a egative photoresist, DNR-L300 (Tokyo electronic materials), was mployed for the lift-off process. The principal fabrication sequence or the polymer-based modified glass substrates and the measure- ent setup are illustrated in Fig. 3. After the fabrication sequence, .5 mm \u00d7 1.5 mm hydrophilic domains surrounded by hydrophoic matrices were formed, as illustrated in Fig. 1(b). Afterward, the - l DI water droplets were positioned at the boundaries of the abricated patterns by an automatic syringe, as shown in Fig. 3(b). he movements of the droplets were recorded immediately after ropping by a digital camera built into the analyzer. Then, we comared the relative LPDs for the patterned materials by analyzing heir mean velocities. We also investigated the maximum liquid olume confined in the barrier layers by adding droplets to analyze he relationship between the volume and LPD. .3. Stability of hermeticity The lens-cover glass with heterogeneous surface energy was onded to a lens body fabricated by silicon micromachining techologies [13]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001343_ichr.2007.4813886-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001343_ichr.2007.4813886-Figure4-1.png",
+        "caption": "Fig. 4. 2-D model of HOAP-3",
+        "texts": [],
+        "surrounding_texts": [
+            "The double pendulum (cf. fig.3) is used, in section IV, as a simple model. We define the initial value qi = [0,0] to get the foot position : (x = 0, y = 0). The final joints value is computed to do a step lenght of d : q f = [ Acos ( d 2L ) ,2\u00d7Acos ( d 2L )] (12) Let us consider one parameter per joint. So, in the case of Fig.3, we get 2 parameters : P = [p1; p2]. The goal is to determine the best value of P, which minimizes the energy consumption and guarantees the contraints shown in I-A, thanks to the constrained optimization algorithm. We define the objective function as : F(x) = t f \u2211 t=0 \u2211 i \u03932 i(t) (13) Where \u0393i(t) is the torque of joint i at the instant t:"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001493_isse.2007.4432850-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001493_isse.2007.4432850-Figure1-1.png",
+        "caption": "Fig. 1. Conductimetric pH Sensor",
+        "texts": [
+            " Drop coating deposition ofpolymers A 2 ,ul drop of each PANI/PVB solution was deposited onto the IDE electrode pattern by use of a Transferpette\u00ae (Sigma Aldrich) piston operated pipette. The drop is suspended on the pipette tip and is pulled onto the substrate by surface tension. The substrates were then placed into an oven for 3 hours to facilitate solvent evapouration. This process was repeated four more times to ensure a good polymer coating over the electrodes. A schematic diagram of the resulting sensors can be seen in Fig. 1. 1-4244-1218-8/07/$25.00 \u00a92007 IEEE 214 30th ISSE 2007 2.4. DC andAC characteristic testing Once the films were deposited, the DC and AC characteristics of the sensors were examined to determine the correct operating conditions for the devices. The ES sensors had a baseline resistance of approximately 1.5 kQ and permitted the use of the in house developed IVR profiler. A National Instruments Data Acquisition (DAQ) card controlled by LabWindows/CVI software and driven by customized electronics hardware measured the current voltage characteristics of the sensors",
+            " The water dispersed ES sensors were not tested for AC conductivity due to the poor adhesion of the polymer to the substrate. The polymer was removed from the sensor surface during the current-voltage tests taken at various temperatures. This occurred when the temperature dropped below the dew point, 1-4244-1218-8/07/$25.00 \u00a92007 IEEE 215 30th ISSE 2007 moisture formed on the film surface removing most of the PANI film. Images of the films taken using an optical microscope can be seen in Fig. 4, which also shows how the PVB/ES Fig. 1 (b) sensors have a much higher surface roughness due to the particle size than with the PVB/EB sensors Fig. 1 (c). 3.3. Effect ofpH on PANI Composite Sensors The PVB/ES and PVB/EB composite sensors were tested separately for pH response. Each sensor was exposed to a solution of known pH and the change in conductance was measured. The reaction between the polymer films and the ions in the test buffers was not instantaneous. A settling time of 6 seconds was required for the conductance values to level off. Fig. 5 shows the settling characteristics of both ES and EB films. After each buffer was applied to the sensors and the change in conductance recorded, the conductance values were plotted against pH to observe any possible relationship between the pH of the test buffer and the conductance of the film"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure5-1.png",
+        "caption": "Fig. 5 Generating process using tw erating circular-arc rack: double-con",
+        "texts": [
+            "org/ on 01/28/201 can be solved by Eqs. 11 and 12 . Likewise, the work gear surface equations can be solved based on system equations Eqs. 6 \u2013 12 . The generating cutting process of the longitudinal cycloidal helical gear is commonly explained by an imaginary virtual generating circular-arc rack formed by the locus of the cutter blades. In this process, all inside cutter blades mounted on the head cutter are used for a double-convex gear see Fig. 4 , and all outside cutter blades are used for a double-concave gear see Fig. 5 . 5 Simulation of the Condition of Meshing and Contact The TCA technique has also been applied to simulate the condition of meshing, whose main goal is to determine the shift of the bearing contact and the transmission errors caused by the misalignment of the gear drive. The first step in such application is to represent the equations for two mating tooth surfaces in the same coordinate system, whose coordinates and unit normal should be the same at the point of contact. Then the kinematic characteristics of this type of gearing can be analyzed by combining TCA with the mathematical model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003734_j.engfailanal.2012.11.013-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003734_j.engfailanal.2012.11.013-Figure4-1.png",
+        "caption": "Fig. 4. Position of the specimens machined from the fractured gear.",
+        "texts": [
+            " To determine the stress intensity factor (SIF) for the crack tip the defect was considered much smaller than the gear (semi-infinite body) and type I loading mode was considered. With this consideration the SIF can be described as a function of applied load (r) and defect dimension (a) according to the following equation: K \u00bc r ffiffiffiffiffiffi pa p \u00f01\u00de FCG tests were conducted in accordance with the ASTM E647 [5] test method using compact tension specimens with thickness (B) of 15 mm and width (W) of 30 mm machined from the core of fractured gear. Fig. 4 shows the plane of the gear teeth from where the FCG specimens were machined. Both fatigue pre-cracking and the FCG tests were performed in a MTS 810 servo hydraulic. Load decreasing up to a crack growth rate of 10 9 m/cycle was used to determine the fatigue threshold. For the fatigue crack growth testing a loading ratio R = 0.1 was used. Table 1 shows the chemical composition of the material used to fabricate the reducing gear. After machining, the gear was carburized, quenched and tempered. The final microstructure was tempered high carbon martensite at the case with a mixture of bainite and low carbon martensite at the core (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000021_esda2006-95565-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000021_esda2006-95565-Figure2-1.png",
+        "caption": "FIGURE 2. MODELLING OF A CANTILEVER BEAM BY STANGO AND SHIA [7]",
+        "texts": [
+            " A large deflection mechanics method is proposed by Stango and Shia [7] to model the deformation of the bristles of a freely rotating cup brush for surface finishing applications. This method has been used by other researchers (e.g., [8]). As the bristle is subjected to a displacement-dependent centrifugal force, it is stated that Eq. (1) is not amenable to a direct analytical solution. Thus, discretisation of the bristle is carried out in conjunction with a numerical method. The bristle is modelled as a system that comprises a number of straight, rigid elements connected by rotational springs, as shown in Fig. 2. loaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/03/2016 Te The length of every link is equal to the bristle length, l, divided by the number of elements, n. The ith link (i = 1, 2, 3\u2026n) has two coordinates that are measured with respect to the axis of rotation: the radial coordinate, ri, which is the distance from the centre of mass, and the slope, \u03b8i, which is the slope of the link. The rotational spring constant, k, is determined by analysing a beam subjected to a constant bending moment, as illustrated in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003689_20121023-3-fr-4025.00027-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003689_20121023-3-fr-4025.00027-Figure2-1.png",
+        "caption": "Fig. 2. Forces acting on the vehicle body.",
+        "texts": [
+            " The load inertia consists of the moments of inertia of the differential gear, the wheel shafts and the wheels. The angular speeds of motor and load are denoted by \u03c9m and \u03c9l, where the load speed is the mean value of the rear wheel speeds, i.e. \u03c9l = 1 2 (\u03c9r,l + \u03c9r,r) . (2) Tm and Tl represent the motor and the load torque, the latter is regarded as an external disturbance. According to Kiencke and Nielsen (2005) the load force Tl/r consists of the mass inertia mv\u0307x, the rolling resistance Fr, the air drag force Fd and the portion of the gravitational force due to the road gradient Fg sin \u03b7, see Fig. 2. The damping coefficient dm considers the viscous friction of the motor. The linear part of the load torque with respect to the load speed is considered separately in the damping coefficient dl. The gear ratio of the axle differential is represented by kg and \u03b8 denotes the total twist angle of the drive line including backlash. The shaft torque Ts(\u03b8) is calculated via a dead zone model as shown in Fig. 3, where the backlash width is 2\u03b1, i.e. Ts(\u03b8) = { ks (\u03b8 \u2212 \u03b1 sgn(\u03b8)) for |\u03b8| > \u03b1 0 for |\u03b8| \u2264 \u03b1 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000200_j.jsv.2006.01.011-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000200_j.jsv.2006.01.011-Figure1-1.png",
+        "caption": "Fig. 1. (a) A uniform beam carrying a 2-dof spring\u2013mass system and (b) replacing the 2-dof spring\u2013mass system is by two equivalent springs keff,i and keff,k.",
+        "texts": [
+            " Four boundary conditions are studied: clamped\u2013free, clamped\u2013simply supported, clamped\u2013clamped, and simply supported\u2013simply supported. It has been found that the agreement between the present results and the FEM results is good. For convenience, the uniform beam with prescribed boundary conditions is called the \u2018\u2018bare\u2019\u2019 beam if it carries no attachment and is called the \u2018\u2018constrained\u2019\u2019 beam if it carries any attachments. 2. Replacing a 2-dof spring\u2013mass system by two equivalent springs For the 2-dof spring\u2013mass system shown in Fig. 1(a), me and Je, respectively, represent the lumped mass and mass moment of inertia of the system, k1 and k2 are the spring constants of the springs, uw and yw are the ARTICLE IN PRESS D.-W. Chen / Journal of Sound and Vibration 295 (2006) 342\u2013361344 translational and rotational displacements of the lumped mass me, a1 and a2 are distances between the center of gravity (c.g.) of the lumped mass and the two springs, ui and yi are the transverse displacement and rotational angle of the uniform beam at the attaching node i, and uk and yk are the those at the attaching node k"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000004_pes.2004.1373063-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000004_pes.2004.1373063-Figure1-1.png",
+        "caption": "Fig. 1. Rotor structure of a buried-type IPM bearingless motor. PM",
+        "texts": [
+            " In this paper, an on-line identification method of the suspension force and the magnetic unbalance pull force parameters is proposed. It is possible that the parameters are accurately identified by a simple experiment. The obtained parameters by the proposed method are compared with the FEM results and discussed. It is confirmed by a proto-type machine that the parameters are accurately identified. The proposed on-line identification method is quite good to realize stable magnetic suspension for an IPM bearingless motor. Fig. 1 shows the buried-type IPM rotor structure. The small permanent magnets are buried just below the rotor core surface. The thin permanent magnets are employed to increase the suspension force for unity MMF of the suspension winding at no load and under light loaded conditions. There are some advantages of buried-type IPM rotor as follows, (1) The airgap length can be small by the buried permanent magnets. Thus, the suspension force for unity MMF of suspension winding is larger than those of surface mounted and inset types"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000070_978-1-4020-4941-5_42-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000070_978-1-4020-4941-5_42-Figure2-1.png",
+        "caption": "Figure 2. Family-1. Figure 3. Family-2. Figure 4. Family-3.",
+        "texts": [
+            ",I JQ Q Q Q to define a linear variety with rank 4, at least one leg for each type is needed, i.e. 1I and . 1J That is, in a feasible 2-dof spherical fully parallel mechanism with legs of US-type, the axes of all the legs must define a degenerate congruence, i.e. the variety of lines which lie in the plane defined by unit vectors k0 and i1 or pass through the point C of that plane. In practice, depending on the varieties of lines spanned by the axes of the legs within a type, three families of mechanism architectures can be identified: Family-1 (Fig. 2): The axes of the legs in the set 1 1 1,..., IQ Q ,..., define a linear variety with rank 1, i.e. a single line passing through C but with direction different to k0, and the axes of the legs in the set 1 2 2 JQ Q ,..., define a linear variety of lines with rank 3, i.e. a plane of lines defined by k0 and i1. Family-2 (Fig. 3): The axes of the legs in the set 1 1 1 IQ Q define a linear variety with rank 2, i.e. a planar pencil of lines with center in R. Vertechy and V. Parenti-Castelli 390 Moreover, addition of legs of type-1 and/or of type-2 to such basic USPMs does not alter the mechanism kinematics but renders the systems redundant and with self-motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003080_jmer.9000033-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003080_jmer.9000033-Figure6-1.png",
+        "caption": "Figure 6. Force diagram of crankshaft.",
+        "texts": [
+            "G: nQQnacy )(sin())2cos()(cos((( 2 54 \u03b8\u03c9\u03b8\u03b8\u03b1 +\u2212\u2212\u2212= nQQ r ))2sin()cos( 2 1 ))2cos()( 23 54 \u03b8\u03b8\u03c9\u03b8 \u2212\u2212 nQQe )((cos))2cos(( 222 23 \u03b8\u03c9\u03b8 \u2212+ jeQQQ r ))sin())2cos()( 1 2 54 \u03b8\u03b8\u2212 (18) Kinetic analysis of slider-crank mechanism Kinetic calculation must start from the piston because slider-crank mechanism started from that. The force diagram of piston was shown in Figure 4 and Equations 19 and 20: \u2211 = PPx amF . (19) ppgx amFR .\u2212= (20) The force diagram of connecting rod was shown in Figure 5 and Equations 21 and 22:     =\u2212 =\u2211 cxcxx cxcx amRN amF . (21) cxcppxx amamRN .. ++= (22) Engine torque can be obtained from Figure 6 as follow: )cos(..)sin(.. \u03b8\u03b8 rNrNT yx += (23) For Ny: \u2211 = \u03b7.AA IM (24) )2cos(. .)sin(.. 23 \u03b8 \u03b7\u03b8 QQ IrN N Ax y + \u2212 = (25) Where IA is Inertia of connection rod. But for all journals: 4321 TTTTTC +++= (26) That Tc indicates the crankshaft torque, not engine output torque. Note that the friction force is negligible in comparison with gas force, so it has ignored in calculations. Engine output torque from flywheel has been calculated considering flywheel inertia and resistance torque of crankshaft end side (Ts)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000016_6.2005-6088-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000016_6.2005-6088-Figure8-1.png",
+        "caption": "Figure 8. Definition of Thruster Force for Controlling Spin Angular Velocity and Out-of-Plane Motion",
+        "texts": [
+            " Because the system rotates around the center of mass of the whole system under the condition that spacecraft are connected by tethers, the system can control the spin radius using only tether tension and centrifugal force while the system must use thrusters to control out-of-plane motion, spin angular velocity and position of the center of mass as shown in Fig.7. This is an especially important feature of the formation flying. Control of thrusters, wheels and tethers is explained in the following parts. Thrusters for controlling spin angular velocity and out-of-plane motion are considered in the coordinate frame V as shown in Fig.8. Let cmr represent a position of the center of the mass of the real formation as follows. cm T cm r}{vr = (41) \u2211\u2211 == = 3 1 3 1 j j j jjcm mrmr (42) Let cmr \u2032 and jr\u2032 represent projections of cmr and jr onto the 21vv plane, and jr\u0302 is defined as follows. cmjj rrr \u2032\u2212\u2032=\u02c6 (43) \u03b8 indicates the angle between d jr and jr\u0302 , and 3jr represents the out-of-plane displacement of spacecraft j . jr\u2206 is defined as shown in the following. j d jj rrr \u02c6\u2212=\u2206 (44) American Institute of Aeronautics and Astronautics 9 Here, jr\u0302 and jr\u2206 are the vectors on the 21vv plan"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure2.14-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure2.14-1.png",
+        "caption": "Figure 2.14 Aircraft flight control PCU schematic",
+        "texts": [
+            " \u2022 Similarly, a system with three integrators will have zero error with a constant acceleration input command. With each integrator in the loop comes a quarter cycle delay at all frequencies. Two integrators would incur a half cycle delay around the loop equivalent to 180 degrees of phase lag. Therefore, even though the integration process has major benefits to control systems by eliminating steady state and even dynamic errors, there are attendant stability issues that must be dealt with in the control system design process. Let us now consider a more sophisticated example. Figure 2.14 shows a schematic of a typical power control unit (PCU) for an aircraft flight control surface. The input end of the summation link is connected via cables and pulleys to the pilot\u2019s control column. This hydraulic position servo actuator allows the pilot to overcome the large hinge moments that occur during high speed flight. 38 Closing the Loop Movement of the mechanical input xi moves the spool valve to the left causing hydraulic fluid to flow from the supply pressure source PS into the piston chamber to the left of the piston"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002169_iros.2009.5354046-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002169_iros.2009.5354046-Figure7-1.png",
+        "caption": "Fig. 7. Disassembled unit",
+        "texts": [
+            " If there exists a wall that can guide the head unit, the guide rollers will suffice for exploration. Fig. 6 shows the front view of the prototype. The maximum bending angle of each joint is about 14 degrees. The center distance of each joint is 16 (mm), so the minimal inner radius is approximately 100 (mm). The head unit has guide rollers on the both sides, so no wire is installed to manipulate it. The materials consists of aluminum alloy (duralumin), brass, and polyacetal. The disassembled components are shown in Fig. 7. All parts are machined using CNC. Gear cogs are not fully made to avoid interference between coaxially located gears A and B. High pressure air is supplied from the last unit, i.e., the tail, and the line is succeeded via silicone flexible tube to the antecedent units. Each unit weighs 31.5 (g) including tubes. Table I summarizes brief specifications of the components. Each joint produces a torque up to 0.042 (Nm) under 0.1 (MPa) of pressure supply. Although air is used in the following cases, liquids including water or oil are available as long as the fluid servomechanism works"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002683_j.finel.2010.08.001-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002683_j.finel.2010.08.001-Figure10-1.png",
+        "caption": "Fig. 10. Quadrilateral and hexahedral",
+        "texts": [],
+        "surrounding_texts": [
+            "In this section, after giving the definition of the involved energetic quantities, we describe the system of the statical and dynamical motion equations and the corresponding adopted solution schemes. elements: topological definition."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000624_00405000701592975-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000624_00405000701592975-Figure2-1.png",
+        "caption": "Figure 2. Deformation of element of fibre due to applied force Cj .",
+        "texts": [
+            " (9) Thermal bonded structure may be assumed as a series of fibre segments bonded by two bonding points. In other words, the fibre segment can be treated as a beam with the built-in ends. According to Pan et al. (1997), the forces acting on each fibre segment can be divided into normal or tangential components. The normal component (Cjn) tends to bend the fibre midway between the bonds whereas tangential component (Cjp) stretches the bond and fibre segment along the fibre axis, as illustrated in Figure 2. In our analysis, the tangential component is neglected; therefore, the deflection of fibre segment with built-in ends has been considered, as shown below (Pan et al., 1997): \u03b4i bend = Cjn(mlb\u0304)3 3Ef iIf i (10) where \u03b4i bend is the bending deflection of the type i fibre segment, Cjn is the normal component of the external load, ml represents the proportion of fibre segment length between the two bonds, Ef i and If i are the elastic modulus and moment of inertia of fibre type i. Also, ml = 1 \u2212 8Vf IR \u03c0 (11) where R = \u03c0\u222b 0 d\u03b8 \u03c0\u222b 0 (\u03b8, \u03d5)J \u2032(\u03b8, \u03d5) sin \u03b8d\u03d5 (12) and J \u2032(\u03b8, \u03d5) = \u03b8 \u2032 2\u222b \u03b8 \u2032 1 d\u03b8 \u03d5\u2032 2\u222b \u03d5\u2032 1 (\u03b8 \u2032, \u03d5\u2032) sin \u03b8 \u2032 1 sin \u03c7 d\u03d5\u2032 (13) \u03c0 \u2212 sin\u22121 ( 1 s ) > \u03c7 > sin\u22121 ( 1 s ) (14) where s is the fibre aspect ratio"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002990_iros.2010.5649131-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002990_iros.2010.5649131-Figure1-1.png",
+        "caption": "Fig. 1. Dynamics model of satellite capture",
+        "texts": [
+            " By using the concept, the appropriate impedance condition to prevent the pushing of the target away after the initial contact is discussed. In this chapter, a free-flying space robot (hereinafter called \u201cchaser\u201d) is assumed to consist of a rigid main body and an n DOF serial link manipulator. The target satellite is also a rigid body. The chaser has already completed the rendezvous with the target. In other to capture the target, the chaser\u2019s manipulator makes contact with the target surface. Figure 1 shows the image of the system of satellite capture. The dynamic equations of a free-flying robot with a manipulator arm are given as follows: [Fb \u03c4 ] = [ Hb Hbm HT bm Hm ] [ x\u0308b \u03c6\u0308 ] + [ cb cm ] \u2212 [ JT b JT m ] Fh, (1) where 978-1-4244-6676-4/10/$25.00 \u00a92010 IEEE 4101 Hb \u2208 R6\u00d76 : inertia matrix of the base Hm \u2208 Rn\u00d7n : inertia matrix of the manipulator Hbm \u2208 R6\u00d7n : coupling inertia matrix between the base and manipulator cb \u2208 R6 : non-linear velocity dependent term of the base cm \u2208 R6 : non-linear velocity dependent term of the manipulator Jb \u2208 R6\u00d7n : Jacobian matrix between the base and end tip of manipulator Jm \u2208 R6\u00d7n : Jacobian matrix between the joints and end tip of manipulator Fb \u2208 R6 : external force and moment on the gravity center of the base Fh \u2208 R6 : external force and moment on the end tip of the manipulator \u03c4 \u2208 Rn : joint torque of the manipulator xb \u2208 R6 : position of the base \u03c6 \u2208 Rn : joint angles of the manipulator The general form of the contact force can be formulated as follows [8]: F = d\u03b4p\u03b4\u0307q + k\u03b4n, (2) where \u03b4 is the penetration"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000909_s12204-009-0681-3-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000909_s12204-009-0681-3-Figure1-1.png",
+        "caption": "Fig. 1 Schematic drawing of a SAW tag",
+        "texts": [
+            " Although they cannot be distinguished by time domain separation directly, the collided tags have special features in Walsh transform (WT) domain. When the feature of a certain tag is matched to the Walsh sequency of the collided multiple tags, the tag is then identified. Comparing with the autocorrelation matched filter, the Walsh threshold matched-filtering is particularly attractive when real-time processing is carried out and/or large amounts of collision data are to be handled. The experimental results manifest the validity of this algorithm. A schematic drawing of a SAW tag is shown in Fig. 1. A radio frequency (RF) pulse is emitted from reader which is basically a radar unit. The RF pulse is received by the antenna of the SAW tag. The interdigital transducer (IDT) is connected to the tag antenna and transforms the received signal into a SAW which propagates along the piezoelectric crystal and is partially reflected by reflectors placed in the acoustic path. The reflected waves are reconverted into a train of electromagnetic pulses by the IDT and are then retransmitted to reader. The IDT behaves like a reciprocal and linear time invariant (LTI) system with an electrical and acoustic gate and can be described with its electroacoustic impulse response hIDT(t)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002777_cdc.2012.6426745-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002777_cdc.2012.6426745-Figure1-1.png",
+        "caption": "Fig. 1. Flexible V-shaped formation during turning maneuver",
+        "texts": [
+            "00 2012 I In this work, we consider the formation motion planning problem for flexible formation of fixed-wing UAVs. In contrast to rigid formation schemes where the controllers aim to maintain constant rectilinear separations between the vehicles, flexible formation scheme allows the geometrical distances between the UAVs to vary slightly when the formation is executing a turning manoeuvre even though the relative curvilinear coordinates that define the formation configuration remains unchanged. Figure 1 depicts a Vshaped formation configuration when it is traveling along a nonlinear virtual leader trajectory, where Figure 2 depicts the same formation configuration when it is traveling along a linear virtual leader trajectory. Here we called the formation virtual leader trajectory as the formation group trajectory. Formation group trajectory represents the trajectory that the formation is planned to travel and it is assumed that it satisfies the nonholonomic kinematic constraint. The virtual structure centre point (virtual leader\u2019s position) is the reference point of the desired formation where the reference position of each UAV Ui of the formation is defined by a relative curvilinear coordinate (pi, qi) with respect to the virtual leader\u2019s position. In this work, (x, y, \u03c8) denotes the pose of the virtual leader (see Figure 1 and Figure 2). This definition of UAV formation position enables us to compute a curve compliant formation reference trajectory that satisfies the nonholonomic kinematic constraint of each UAV if the formation group trajectory satisfies the nonholonomic constraint [1]. When every UAV executes its respective formation reference trajectory stably, the flexible formation motion control is achieved and maintained along the planned formation trajectory. In this work, the obstacles or the no-fly zones (NFZ) in the environment are assumed to be static and can be modeled by obstacle circles"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000772_robio.2007.4522451-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000772_robio.2007.4522451-Figure5-1.png",
+        "caption": "Fig. 5 Fluid analyses inside the suction",
+        "texts": [
+            " 3r 4r - Circumference\u2019s inside radius and outside radius of the sealed ring (m). Because of t ap p and 4 3r r , 0Q . It accords with the real airflow direction that enters into the cup from outside. According to the formula (11) and 4 3r r r , the relation is simulated, see Fig. 4. From Fig.4, Q increases with the gap h, but it reduces when r , the thickness of the sealed ring increases. This supplies reference for the robot structure. When the airflow is steady in the suction cup, take on the control system, it is between the section and , see Fig. 5. 1v 1p 1A and 2v 2p 2A are respectively its airflow velocity, pressure and area of section, and MQ is mass airflow. To section and , the continuity equation is: 1 1 2 2v A v A According to fluid network theory [10], there is much damp because of the different section area and the part air loss in the air passage: 1 22 2 1 2 , 2 2 M MQ QR R A A , 2 22 MQR A then 1 2 2 1( )Mp p Q R R R 12 2 1R R R R 12R - Equivalent damp between section and (1/m.s). Unless pipeline length l d pipeline diameter , because the damp in this case can\u2019t be calculated using the method in the pipe case directly, so we can use this method above to calculate approximatively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002311_978-3-642-17390-5_19-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002311_978-3-642-17390-5_19-Figure9-1.png",
+        "caption": "Fig. 9. Various hardware modules used by the Kinbots to implement the GSO behaviors",
+        "texts": [
+            " Certain algorithmic aspects need modifications while implementing in a robotic network mainly because of the point-agent model of the basic GSO algorithm and the physical dimensions and dynamics of a real robot. We used a set of four wheeled robots, called Kinbots (See Figure 8), that were originally built for experiments related to robot formations [23]. By making necessary modifications to the Kinbot hardware, the robots are endowed with the capabilities required to implement the various behavioral primitives of GSO. The various hardware modules of each robot that are used to achieve the above tasks are shown in Figure 9. We presented the results from an experiment where two Kinbots use GSO to localize a light source. The paths traced by the robots as they execute the GSO is shown in Figure 10. Kinbots implementing GSO to localize a sound source is demonstrated in [24]. In [25], we investigated the behavior of agents that implement GSO when mobile sources are considered. In particular, we use GSO to develop a coordination scheme that enables a swarm of mobile pursuers, with hard-limited sensing ranges, to split into subgroups, exhibit simultaneous taxis toward, and eventually pursue a group of mobile signal sources"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001699_s1068798x10100072-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001699_s1068798x10100072-Figure1-1.png",
+        "caption": "Fig. 1. Model of transmission with intermediate rollers and a free iron ring.",
+        "texts": [
+            "001 Transmissions with intermediate rollers has techni cal data in complex: for example, compactness, high output torque, a large single stage gear ratio (2\u201350), increased load capacity, low inertia, small gaps in engagement, and long life. Accordingly, the industrial application of transmissions with intermediate rollers continues to expand (for example, in the oil and gas industry, in transport systems and cranes, in medical and metallurgical equipment, and in the aerospace industry). The best such system is a transmission with inter mediate rollers and a free iron ring (Fig. 1). With cer tain parameter combinations, sliding friction at engagement is close to zero. (Rolling friction predom inates.) Hence, high efficiency is possible. In addition, this transmission exhibits multipair engagement, and consequently brief overloads may be withstood with out injury. On the basis of this transmission, it is expe dient to design planetary mechanisms, since the gears in mutual engagement have little difference in the number of teeth. Despite the wide use of existing transmissions with intermediate rollers and the appearance of new designs, there has been little study of such systems"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002323_978-3-642-25486-4_25-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002323_978-3-642-25486-4_25-Figure6-1.png",
+        "caption": "Fig. 6. Elasticity contributors",
+        "texts": [],
+        "surrounding_texts": [
+            "1. Merlet, J.-P.: Parallel Robo 2. Stewart, D.: A Platform Engineers Proceedings, Vo 3. Clavel, R.: Device for the No. 4976582 (December 1 4. Clavel, R.: Delta, a fast ro Symposium on Industrial R 5. Tsai, L.-W.: Robot Analy Wiley & Sons (1999) 978-0 6. Nefzi, M.: Kinematics an Vector-based kinematic cal 7. Gosselin, C.: Stiffness Ma and Automation 6(3), 377\u2013 8. Timoshenko, S.: Strength & Distributors (2004) 978- Stiffness Analysis of Clavel\u2019s DELTA Robot to maximum deflection for reference load in X- and Y-directi tors to maximum deflection for reference load in Z-direction ts. Springer, Heidelberg (2006), 978-1402041327 with Six Degrees of Freedom, UK Institution of Mechan l. 180(15), Pt. 1 (1965) Movement and Positioning of an Element in Space, US Pa 1, 1990) bot with parallel geometry. In: Proceedings of 18th Internatio obots, pp. 91\u2013100 (1988) sis: The Mechanics of Serial and Parallel Manipulators. J 471325932 d Dynamics of Robots, Exercise RWTH Aachen, Germa culations for parallel robots, not published (2010) pping for Parallel Manipulators. IEEETransactions on Robo 382 (1990) of Materials: Elementary Theory and Problems. CBS Publish 8123910307 249 on ical tent nal ohn ny: tics ers"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000393_9780470061565.hbb032-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000393_9780470061565.hbb032-Figure8-1.png",
+        "caption": "Figure 8. Nyquist plot with impedance vector.",
+        "texts": [
+            " Using Euler\u2019s relationship exp(j\u03c6) = cos \u03c6 + jsin \u03c6 (22) to express the impedance as a complex function, the potential is described as, E(t) = E0 exp(j\u03c9t) (23) and the current response as, I (t) = I0 exp(j\u03c9t \u2212 j\u03c6) (24) The impedance is then represented as a complex number, Z = E I = Z0 exp(j\u03c6) = Z0(cos \u03c6 + j sin \u03c6) (25) The expression for Z(\u03c9) is composed of a real and an imaginary part. Plotting the real part on the horizontal axis and the imaginary part on the vertical axis results in the \u201cNyquist plot\u201d. In this plot, the vertical axis is negative and each point on the Nyquist plot is the impedance at one frequency. In Figure 8 low-frequency data are on the right side of the plot and higher frequencies are on the left. This is normally true for EIS data where impedance usually falls as frequency rises but not true for all circuits. The major drawback of Nyquist plots lies in the fact that looking at any data point on the plot, one cannot tell what frequency was used to record that point. The Nyquist plot in Figure 8 results from the electrical circuit shown in Figure 9 with Rs = 0. The semicircle is characteristic of a single \u201ctime constant\u201d. Electrochemical impedance plots often contain several time constants. Often only a portion of one or more of their semicircles is seen. EIS data is commonly analyzed by fitting it to an equivalent electrical circuit model. Most of the circuit elements in the model are common electrical elements such as resistors, capacitors, and inductors. To be useful, the elements in the model should have a basis in the physical electrochemistry of the system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003205_peoco.2012.6230916-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003205_peoco.2012.6230916-Figure1-1.png",
+        "caption": "Fig. 1. Excitation characteristics, excitation current and flux of a power transformer with and without dc bias",
+        "texts": [
+            " Consequently, the core is saturated Effect of DC Bias on Magnetization Current Waveforms of Single Phase Power Transformer Syafruddin Hasan, Soib Taib, Risnidar Chan, Siti R.M N 978-1-4673-0662-1/12/$31.00 \u00a92012 IEEE 496 during the half cycle in which the bias current is in the same direction as the magnetizing current [2]. This additional flux \"bias\" or \"offset\" will push the alternating flux waveform closer to saturation in the positive cycle then the negative cycle. Therefore, the core material is said to experience half-cycle saturation. The core material will enter half-cycle saturation earlier if the magnitude of DC component is greater. Fig.1 displays the excitation characteristics, exciting current and flux of a power transformer iron core. Explanation of Fig.1, ac = ac flux produced by ac current dc = dc flux produced by dc current = ac + dc = total flux in the iron core s = saturation flux = the minimum angle in one cycle at which = s M = knee point N = saturation point Is = saturation current Idc = direct current due to dc bias k1 and k2 are gradients of piecewise magnetizing characteristic It can be seen that the exciting current is symmetrical to time axis without dc bias, and it becomes asymmetrical with dc bias occurring. Under such condition, the dc flux generated by the dc current offsets the ac flux in the transformer iron core. This is the so called half-cycle saturation of transformer under dc bias [9]. Since a typical power transformer needs little magnetizing current that set to the knee point at rated condition, a small amount of direct current will sufficient apt to cause significant half-cycle saturation. From Fig.1, we can found that the exciting current waveform with dc bias is sharper than without dc bias. Refer to flux or magnetizing current waveforms that exhibit asymmetrical characteristics, i.e., i(t) \u2260 - i(-t), the even harmonics will produced significantly. In this investigation, the magnetizing (exciting) current prediction program is developed by using MATLAB software. The concept of the program is to return a graphical approximation of the B-H relationship when a transformer\u2019s core magnetizing characteristic is known and a certain level of DC bias is assumed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003653_phm.2012.6228866-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003653_phm.2012.6228866-Figure5-1.png",
+        "caption": "Figure 5. Broken tooth width of gear 2.",
+        "texts": [
+            " Gears land 2 are the tested gears. Crack is a very common fault mode studied in gear fault diagnosis. For this reason, crack faults were planted in our gearbox experiments. Let a be the crack angle and the bold line the face width, as shown in Figure 3. The crack depth of gear 1 are Imm and 2mm as depicted in Figure 4. Broken tooth is also a very common fault mode in gear fault diagnostics. Therefore, broken tooth faults were embedded in gear 2. The broken widths are 2mm, 5mm, and IOmm respectively as depicted in Figure 5. As a result, two gears with six different conditions including one normal gear and five faulty gears are tested in the experiment. 978-1-4577-1911-0/12/$26.00 \u00a92012 IEEE MU3035 2012 Prognostics & System Health Management Conference (PHM-2012 Beijing) The vibration was measured for the two gears using four acceleration sensors. They were mounted on the gearbox casing as depicted in Figure 2. Labview software was used to collect the vibration data for further processing. The speed of the driving motor and the load of the magnetic brake vary to simulate the general gearbox operating conditions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure9.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure9.4-1.png",
+        "caption": "Fig. 9.4-1 Torsional vibration of a long rod.",
+        "texts": [
+            " Specifi cally, since it is physically diffi cult to VibrationAnalysis_txt.indb 319 11/24/10 11:49:43 AM Principles of Vibration Analysis 320 take measurements in the recessed space, we can simply attach a slender bar to the desired measurement location and then measure the accelerance H11 at the free end of the bar. Th en, using Eq. l, once H11 is known, the accelerance H22 of the structure is readily determined. 9.4 Torsional Vibration of Rods Consider next the twisting oscillation of a circular rod, as represented in Fig. 9.4-1 [9.2, 9.3]. Let be the angle of twist. Th en, will depend upon the axial position x along the rod, and upon the time t. Th at is, = ( , )x t . Consider a diff erential element with length dx at location x along the rod, as represented in Fig. 9.4-1. From Hooke\u2019s law, the axial moment T at x is T J G xp= \u2202 \u2202 , (9.4-1) where Jp is the second polar moment of area of the cross section and G is the shear modulus. Th e product JpG is known as the torsional stiff ness of the rod. Applying Newton\u2019s law with a free-body diagram of the element, we have J t dx T T x dx Tp \u2202 \u2202 = + \u2202 \u2202 \u239b \u239d\u239c \u239e \u23a0\u239f \u2212 2 2 , (9.4-2) where \u03c1 is the mass density (mass per unit volume). By substituting for T from Eq. 9.4-1 into Eq. 9.4-2, we have the equation of motion J t x J G xp p \u2202 \u2202 = \u2202 \u2202 \u2202 \u2202 \u239b \u239d\u239c \u239e \u23a0\u239f 2 2 (9",
+            "indb 320 11/24/10 11:49:44 AM Continuous Systems | Chapter 9 321 Eq. 9.4-4 is identical in form to Eq. 9.2-3 (the wave equation for the vibrating string). Th en, by comparison, following the same solution procedure, we obtain in the form ( , ) sin cos sin cos ,x t A G x B G x C t D t= + \u239b \u239d \u239c \u239e \u23a0 \u239f +( ) (9.4-5) where, as before, the coeffi cients A, B, C, and D are to be determined by the boundary and initial conditions. Table 9.4-1 lists the common boundary conditions for torsional vibration of a rod (see [9.2]). Example 9.4-1 Fig. 9.4-1 depicts a long uniform rod with one end fi xed and the other free. Suppose a torque T is applied at the free end and then released. Find the subsequent torsional motion. Solution: From Eq. 9.4-5, the general solution has the form ( , ) sin cos sin cos ,x t A G x B G x C t D t= + \u239b \u239d \u239c \u239e \u23a0 \u239f +( ) (a) VibrationAnalysis_txt.indb 321 11/24/10 11:49:45 AM Principles of Vibration Analysis 322 where is mass per unit volume, and G is the shear modulus. Th e boundary conditions are ( , )0 0t = (b) \u2202 \u2202 = = x x L 0 (c) Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001225_j.cnsns.2009.02.028-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001225_j.cnsns.2009.02.028-Figure1-1.png",
+        "caption": "Fig. 1. Relative coordinates with vehicle i\u00fe 1 pursuing vehicle i.",
+        "texts": [
+            " tj and xj are the control inputs (linear velocity and angular velocity). The vehicle with index 0 will be referred to as the leader, and the others as followers. Our focus here is to analyze the leader\u2013follower formation, when the trajectory of the leader is a straight line or a circle. These trajectories for the leader are obtained by the simple control law t0 \u00bc V ; x0 \u00bc x: \u00f02\u00de The case x \u00bc 0 represents straight line motion, while x\u20130 corresponds to circular motion. The configuration of the n-unicycle system is shown in Fig. 1, where ri is the relative distance between the two vehicles, ai is the angle between the current orientation of the ith unicycle and the line of sight, and bi is the angle between the current orientation of i\u00fe 1th unicycle and the line of sight. Both angles are positive in the sense of counterclockwise rotation to the line of sight. Following [14], the kinematic equations are written in relative coordinates: _r0 \u00bc t0 cos a0 t1 cos b0; _a0 \u00bc 1 r0 \u00f0t0 sina0 \u00fe t1 sin b0\u00de x0; _b0 \u00bc 1 r0 \u00f0t0 sina0 \u00fe t1 sin b0\u00de x1; 8>< >: _ri \u00bc ti cos ai ti\u00fe1 cos bi; _ai \u00bc 1 ri \u00f0ti sinai \u00fe ti\u00fe1 sin bi\u00de xi; _bi \u00bc 1 ri \u00f0ti sinai \u00fe ti\u00fe1 sin bi\u00de xi\u00fe1; 8>>< >: \u00f03\u00de where i \u00bc 1; "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003982_clen.201200075-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003982_clen.201200075-Figure3-1.png",
+        "caption": "Figure 3. Conductivity removal efficiencies of the reverse osmosis membranes.",
+        "texts": [
+            "7% [35], for treating the wastewaters of the leather industry. Accordingly, the COD removal efficiencies obtained with the BW30 and XLE membranes in this study are considered to be satisfactory. Also, with regard to these short-term studies, the implementation of ultrafiltration or the failure to do so did not cause any significant differences in the removal efficiencies. Conductivity was measured in the filtered waters obtained after the application of the reverse osmosis process, and the achieved conductivity removal efficiencies are shown in Fig. 3. The conductivity removal efficiencies yielded similar results in comparison to the COD removal efficiencies. Similarly, the reverse osmosis process carried out without ultrafiltration showed somewhat higher amounts of conductivity removal. The root cause of this is believed to be the fact that, as explained above, there is an increase in the efficiency of retaining dissolved material due to the plugging of the membrane and the formation of a foulant layer. Similar to the COD removal, the highest conductivity removal efficiencies were Table 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001109_s11740-008-0080-x-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001109_s11740-008-0080-x-Figure6-1.png",
+        "caption": "Fig. 6 Cross sectional area with respect to the powder feed rate and the modulation frequency",
+        "texts": [
+            " The penetration depth with respect to the modulation frequency and the powder feed rate is given in Fig. 5. The penetration depth is varying between 0.5 and 2.7 mm. The largest penetration depth was measured at a modulation frequency of 10 Hz and a powder feed rate of 3.6 g/min. For the powder feed rate of 3.6 g/min the clear behavior of the depth with respect to the modulation frequency is shown. Beside the penetration depth of the track the cross sectional area of the tracks was measured. The cross section versus the process parameters is given in Fig. 6. The smallest area is 1.3 mm2 and the largest area is 6 mm2. The largest cross sectional area was measured at 10 Hz modulation frequency and 3.6 g/min powder feed rate. The parameter combination, which shows the largest area shows the largest depth. The dependency of the area on the process parameters is in good correlation to the depth. In general the area is decreasing with an increasing modulation frequency. The cross sectional area can be taken as a value for the energetic efficiency. The area is proportional to the treated volume per energy"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003075_icra.2011.5979608-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003075_icra.2011.5979608-Figure6-1.png",
+        "caption": "Fig. 6. Computing the height of a blob using the camera geometry and HRP-2 kinematics.",
+        "texts": [
+            " For more robust recognition, we can apply the threshold of the height of a knob from the floor, Th. In general the knob should be located in the reachable height not only for the normal people but also for the handicapped. Since the robot knows the transformation matrix from the ground coordinate, \u03a3G that is attached under the left foot, to the coordinate of the left camera, \u03a3C , it can compute the height of the blob by using the simple geometry of the view-angle in vertical, \u03b8 and the distance to the blob, d as shown in Fig. 6. Note that the camera of the robot should be parallel to the floor plane to compute the height of a knob in the following method. The d can be obtained by subtracting the walked distance from the door distance at the initial location. To obtain the ratio \u03ba of length per pixel, h in the figure is computed, followed by the height of the knob hknob from the floor by adding the height of the camera, hcam as follows: h = d\u00d7 tan(\u03b8/2), (1) \u03ba = h Iy \u2212 Pcy , (2) hknob = \u03ba(Pcy \u2212 Py) + hcam (3) where the Pcy is the y pixel value of the principal point of the camera and Py is that of the centroid of the knob based on the image coordinate, \u03a3I "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000691_icbbe.2008.749-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000691_icbbe.2008.749-Figure1-1.png",
+        "caption": "Figure 1. 3D FEM wire mesh model",
+        "texts": [
+            " In this research, needle works as a rigid model. After an analysis using a full size model, it is found that the stresses far away from the needle insertion point can be negligible. Therefore, the model is constrained so that it can imitate the experimental setup. The edges of back plane are fixed in all directions. Symmetric boundary condition constraint is added on the symmetrical plane. Other faces and edges are left free to distort. Node rendered in black point works as the target point. The wire model is shown in Figure 1. Figure 1 (a) displays the 0=t global mesh of prostate sample and the needle model. Figure 1(b) shows itt = model deformation during the needle inserts into it. Before the surgery, locations of the seed (in this paper we name the planned seed location as target) are available through the optimal plan. Achieving the target is left to the physician. Seeds are ejected into the prostate gland when the depth specified by the plan is reached. Unfortunately, force load on the tip of insertion needle will cause the tissue deformation during the process and result in the misplaced seeds, as demonstrate in Figure 2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003727_j.engfailanal.2011.06.004-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003727_j.engfailanal.2011.06.004-Figure4-1.png",
+        "caption": "Fig. 4. Applied forces on the cooling fan assembly.",
+        "texts": [
+            " From input data and detailed results for 9900 ENF 6 fan type, weight, Fw; centrifugal, Fc; axial, Fa; and tangential, Ft; force acting on blades are calculated as follows: Fw \u00bcmg \u00bc \u00f078 kg\u00de 9:81 m s2 \u00bc 765 N \u00f01\u00de Fc \u00bc mrmx2 \u00bc \u00f078 kg\u00de \u00f01:93 m\u00de 10:7 rad s 2 \u00bc 17;235 N \u00f02\u00de Fa \u00bc Axial thrust 6 \u00bc 11;760 N 6 \u00bc 1960 N \u00f03\u00de Ft \u00bc Fan shaft power 6 rA x \u00bc 123;700 W 6 2:58 m 10:7 rad s \u00bc 746 N \u00f04\u00de where m is the mass of each blade and x is the angular velocity of blades. rm is the distance between the origin of the coordinate in the center of the hub plate and mass center. rA is the distance between the origin of the coordinate in the center of the hub plate and area center of blades. After the blade is modeled, rm and rA are calculated using SolidWorks software. The directions of applied forces on blades of the cooling fan are shown in Fig. 4. It is noted that the condition of loading is multiaxial. Among all forces, the only alternating force is weight force. Except the weight, all forces acting on the blades have fixed direction in relative coordinate system whose axes are rotating with the blade. To obtain the stresses of the U-bolts as realistic as possible, first, U-bolts are stretched with a tension of 45 MPa in the normal direction to the hub plate from sections near the nuts. Then, forces are exerted to the blade, which contact with hub plate and U-bolts"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000804_robot.2007.364174-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000804_robot.2007.364174-Figure1-1.png",
+        "caption": "Fig. 1 7 DOF manipulator model.",
+        "texts": [
+            " Finally, section V presents kinematic simulations to show that the proposed method is effective for avoiding the joint limits. 1-4244-0602-1/07/$20.00 \u00a92007 IEEE. 4510 This section provides an analytical inverse kinematic solution for a 7 DOF manipulator model. First, the manipulator model assumed in this paper is described. Next, a parameter is introduced to represent the redundancy. Then, the inverse kinematic solution is derived using the redundancy parameter. In this paper, a S-R-S manipulator model is assumed. Namely, as shown in Fig. 1, the manipulator is assumed to have seven revolute joints, comprising three shoulder joints, one elbow joint, and three wrist joints. To describe the kinematic relation between the joint angles and the pose (position and orientation) of the manipulator\u2019s tip, let us define joint coordinate systems. In this paper, each coordinate system \u03a3i (i = 0, 1, \u00b7 \u00b7 \u00b7 , 7) is determined based on the Denavit-Hartenberg rules [9]. The base coordinate system \u03a30 and the tip one \u03a37 are placed as shown in Fig. 1. With these coordinate systems, the Denavit-Hartenberg parameters are described as listed in Table 1. Note that the notation of the parameters is not unique because the parameters depend on the definition of the joint coordinate frames. Since the tip pose is uniquely described by six parameters, an additional parameter is required to specify the manipulator\u2019s posture uniquely. This parameter is associated with a self-motion of the manipulator. To describe the self-motion, this paper incorporates the parameter termed arm angle [10]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000930_bfb0110318-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000930_bfb0110318-Figure1-1.png",
+        "caption": "Fig. 1. Forces acting on the aircraft.",
+        "texts": [
+            " A simplified model of a VTOL aircraft is given as follows Xl ~ X2 ~ _ sin(Oi) h 2sin(~) = + cos(01) Yi = y2 y2 = - cos(Oi) h + sin(Oi) ~ F - g = f f cos (a )F (18) where M denotes the mass of the aircraft, J the moment of inertia about the center of mass C, l the distance between the wingtips and g the gravitational acceleration. The control inputs are the thrust directed out the bot tom of the aircraft, denoted by T, and the rolling moment produced by the torque F, acting at the wingtips, whose direction with respect to the horizontal body axis is given by some fixed angle a (see figure 1). As expected, xi, center of mass and the roll angle of the aircraft with respect to the horizon, while x2, y~ and 02 the respective velocities. Typica l uncer ta in t ies which the autopi lo t have to deal wi th are given by the value of the mass M (and thus of the momen t of iner t ia J ) and of the angle a . The in ternal model uni t mus t compensa te for both the uncer ta in t ies in the signal to be t racked and the pa ramet r i c uncer ta in t ies of the model . In view of the previous discussion, the control T is assigned as the sum of two te rms T = T~m + Tst represent ing respectively the ou tpu t of the in ternal model uni t and an e x t r a t e rm which, along with F = Fst, is used to g lobal ly s tabi l ize the zero-error manifo ld "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002533_i2011-11121-9-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002533_i2011-11121-9-Figure1-1.png",
+        "caption": "Fig. 1. Geometry of a swimmer.",
+        "texts": [
+            " We investigate this by means of numerical simulations, supported by approximations. There have been a number of recent studies of the motion of pairs of interacting three-sphere and dumb-bell swimmers [27\u201332] and we give comments and comparison in our final discussion section. For the present we simply note that since our swimmers are identical and externally driven, they are naturally synchronised, and so their behaviour may well differ from that of autonomous organisms in biological systems. The basic geometry of a single swimmer is shown in fig. 1. The swimmer is defined by the distance s between the two beads, the angle \u03c6 to some axis and the position X of its centre of reaction, which is the average of the two bead positions weighted by their radii R1 and R2 [33]. Note that we assume the swimmer beads are spherical and are surrounded by fluid in three dimensions; however we take the swimmers to be confined to a plane. In current experiments (not reported here), swimmers sit at a fluidair interface. In dimensional quantities the leading-order equations governing the degrees of freedom (s, \u03c6) are 6\u03c0\u03b7(\u03c72 1R1 + \u03c72 2R2)s\u0307 + k(s \u2212 l0) = 3\u03bc0m1m2 4\u03c0s4 [cos(\u03b12 \u2212 \u03b11) \u2212 3 cos \u03b11 cos \u03b12], (1) 6\u03c0\u03b7(\u03c72 1R1 + \u03c72 2R2)s2\u03c6\u0307 =\u2212Bextb[m1 sin(\u03c6 + \u03b11 \u2212 \u03c8)+m2 sin(\u03c6 + \u03b12 \u2212 \u03c8)]",
+            " To give motion of an individual swimmer, we need to include the leading-order viscous interaction between the moving spherical beads in the system [19,33], resulting in an equation for the centre of reaction X, 6\u03c0\u03b7(R1 + R2)X\u0307 = \u22129\u03c0 2 \u03b7R1R2 s (\u03c72 \u2212 \u03c71)(2s\u0307r\u0302 + s\u03c6\u0307\u03c6\u0302). (3) Here r\u0302 is a unit vector from bead 2 to bead 1 and \u03c6\u0302 is a perpendicular unit vector. For magnetic properties, in this paper we take the simplest soft-hard system used in [21], in which bead 1 is magnetically completely soft, and so its magnetic field direction m1 follows that of the externally imposed magnetic field b(t), with the angle \u03b11 = \u03c8 \u2212 \u03c6 in fig. 1. Bead 2 however is magnetically entirely hard and its field direction m2 is fixed, say along the axis of the swimmer, with direction \u03b12 = 0. In dimensionless variables the degrees of freedom (s, \u03c6) are then governed by the internal dynamics, s\u0307 + s \u2212 1 = \u22122Amags \u22124 cos(\u03c8 \u2212 \u03c6), (4) s2\u03c6\u0307 = Aext\u03c3 \u22121b sin(\u03c8 \u2212 \u03c6), (5) where the time-dependent external magnetic field is given by b \u2261 b(cos \u03c8, sin \u03c8) = (\u03b1 cos t, \u03b2 sin t), (6) with \u03b1 and \u03b2 constant. For a complete discussion of the non-dimensionalisation, we refer the reader to [21]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure10-1.png",
+        "caption": "Fig. 10. Coupled system for global bending (modal approach)",
+        "texts": [
+            " From this modal base, the original FRF's can be synthesized using equation (4) (4) The same FBS approach can now be used but instead of working with the measured FRF's describing the vehicle body subsystem, the synthesized FRF's will be used. Of course a modal truncation error will be made by only taking a limited modal base into account. This effect is very important at the clamping and loading locations. To compensate for the missing flexibility, the residual stiffness at the clamping and loading points can be integrated in the FBS equation. Instead of connecting subsystem A (vehicle body) rigidly to subsystem B (ground), the two subsystems are now connected by a spring representing the residual stiffness as shown in Fig. 10 and 11. This residual stiffness can be estimated based on the upper residual terms obtained from the modal identification of the system under free-free conditions. If one is interested in only the contribution of a subset of modes to the global stiffness, one can limit the FRF synthesis (equation (4)) only to that particular subset of modes. In Fig. 12 the torsional stiffness is plotted over an increasing number of modes taken into account during the FRF synthesis. Large drops indicate modes with a large contribution to the global torsional stiffness"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002125_iciea.2009.5138715-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002125_iciea.2009.5138715-Figure1-1.png",
+        "caption": "Figure 1. Characteristics of 4/3 closed-center proportional valve.",
+        "texts": [
+            " The larger tolerances on spool geometry result in response nonlinearities, especially in the vicinity of neutral spool position. Proportional valves lack the smooth flow properties of \u2018\u2018critical center\u2019\u2019 valves, a condition closely approximated by servo valves at the expense of high machining cost. Small changes in spool geometry (in terms of lapping) may have large effects on the hydraulic system dynamics. For a closed-center spool (overlapped), which usually provides the motion of the actuator in a PHS, may result in steady-state error because of its deadzone characteristics in the flow gain. Fig. 1 illustrates the characteristics of proportional valve with deadzones. Fuzzy logic-based controllers have received considerable interest in recent years [2],[3]. Fuzzy-based methods are useful when precise mathematical formulations are infeasible. Moreover, fuzzy logic controller soften yield superior results to conventional control approaches. However, direct application of usual \u201cfuzzy PD\u201d controllers to a system with deadzones results in poor transient and steady state behavior. In particular, a steady state error occurs when using a fuzzy PD controller to a system with deadzones, the size of the steady state error increases with the dead zone width",
+            "5% Amplifier card set point values \u00b1 10 VDC, solenoid outputs (PWM signal) 24 V, dither frequency 200 Hz, max current 800 mA, DAQ Card NI 6221 PCI analog input resolutions 16 bits (input range \u00b110V), output resolutions 16 bits (output range \u00b110V), 833 kS/s (6 \u03bcs full-scale settling) Operating systems & Program Windows XP, and LabVIEW 8.2 Figure 2. PC-Based position control of a PHS. Consider the system shown in Fig. 3, The transfer function P(z) represents the plant (cylinder), D represents a A proportional vale with deadzones, F[e(k), \u0394e(k)] represents a FLC control law, K1 is the feed forward gain, v(k) is the output of the controller, u(k) is the output of a proportional valve, ym(k) is the reference input, and yp(k) is the output of the plant. The characteristics of the proportional valve with deadzones D (from Fig.1) is described by the function where d, m \u2265 0. The parameter 2d specifies the width of the deadzone, while m represents the slope of the response outside the deadzones. The control scheme of Figure 3 represents a typical FLC control system. The fuzzy PD controller is the controller that uses the error and change of error (proportional-derivative) as inputs, These inputs are similar to a conventional PD controller [3]. The feed forward term K1 is needed in order to eliminate the steady state error [4]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure1-1.png",
+        "caption": "Fig. 1 Profile models on which tooth tip relief is applied: (a) addendum relief, (b) dedendum relief",
+        "texts": [
+            " As a result, the teeth do not perform as they are assumed to in theory. Premature contact at the tips or excessive contact pressures at the end of the teeth give rise to noise and/or gear failures. In order to reduce these causes of excessive tooth load, profile modification is a usual practice [19]. One of these modifications is tooth tip relief. Here, in order to maintain fine gear teeth meshing, a small amount of material is removed from the peak area of the gear teeth. Tooth tip relief is classified into two parts: addendum and dedendum relief (Fig. 1). When comparing the two relief types, addendum relief seems to be widely used. Figure 2 shows Crowning and Side Relief, another type of modification. These two modifications are applied along the axis of the gear teeth surfaces (Fig. 2). Crowning involves removal of a small amount of material from the gear teeth layer, starting from the center towards the edge of the teeth and causing the teeth surfaces to acquire convex-like shapes. Crowning helps to maintain teeth contact in the middle part of the gear teeth as well as preventing edge contact resulting from lower load carrying capacity"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002483_kem.490.97-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002483_kem.490.97-Figure6-1.png",
+        "caption": "Figure 6. Schematic and photograph of Prototype III with top foil removed (see text for symbol definitions).",
+        "texts": [
+            " This finding suggested that the stiffness distribution in bump foil segments (along the angular direction) was not optimal, with the bumps adjacent to the fixed ends being too stiff. Subsequent numerical simulations of the bump foil stiffness (using numerical tool described in [11]) were in agreement with this hypothesis. Prototype III. Based on the experimental observations made while testing Prototype II, it was decided to design and build third prototype, which would have closer to optimal stiffness distribution in the bump foil structure (Figure 6). The design process utilized numerical tools developed earlier in the project [8], [11]. A significantly more uniform stiffness distribution in the bump foils was achieved by varying the radius of curvature of each bump in bump foil segments, with the largest radius at the fixed end and the smallest radius at the free end (Table 2). The number of bumps in a bump strip was also reduced from 8 to 5, because numerical simulations for larger number of bumps predicted that the first bump (near the fixed end) was always significantly stiffer than other bumps due to bump \u201cpinning\u201d caused by frictional forces. This mechanism was described in detail in [11] and [12]. Prototype II consisted of 42 bump foil segments in 7 rows of 6 segments (Figure 6). Experimental results of Prototype III. Figure 7 presents typical experimental results of Prototype III collected at a constant radial load of 3 kN (about 300 kPa) and the two rotational speeds of 800 rpm and 1600 rpm. It can be seen that soon after startup the frictional torque dropped gradually to about 2.2 Nm. Similar reduction in frictional torque was observed in most experimental runs, and although it was not as significant as during the initial run with new top foils (Figure 4), it suggests further wear-in of the top foil during testing"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000924_978-1-4020-5967-4_1-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000924_978-1-4020-5967-4_1-Figure1-1.png",
+        "caption": "FIGURE 1. The screw used with a wooden beam or \u0301 o (Mechanics 2.5). On the left is Drachmann\u2019s drawing (The Mechanical Technology, p. 58) made from Ms B; on the right is the figure from Heronis Alexandrini opera, vol. II, p. 106.",
+        "texts": [
+            " 14 See Heron\u2019s own treatise on the subject, the De automatis, in Opera, vol. I, pp. 338\u2013453. these devices as belonging to a special class, and was quite independent of any theoretical understanding of their operation. Heron begins his account of the five powers with a description of their construction and use (2.1\u20136) that reveals a close familiarity with practitioners\u2019 knowledge.15 The construction of the wheel and axle (2.1), the compound pulley (2.4), and the screw (2.5\u20136) is described in detail. Two uses of the screw are described: with a wooden beam or \u0301 o (Fig. 1), and with a toothed wheel (Fig. 2).16 The account employs a good deal of specialized terminology for the mechanical powers and their parts; a number of these technical terms are explicitly flagged as such using the Greek word \u0302 , \u201cto be called\u201d.17 A striking feature of Heron\u2019s account is the statement of rough, non-quantitative correlations describing the behaviour of the five powers. Thus instead of a precise formulation of the law of the lever as a proportionality between forces and weights, Heron remarks that \u201cthe nearer the fulcrum is to the load, the more easily the weight is moved, as will be explained in the following\u201d (2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003301_r-395-Figure10.5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003301_r-395-Figure10.5-1.png",
+        "caption": "Fig. 10.5-1 An engine mounting system.",
+        "texts": [
+            "indb 379 11/24/10 11:51:01 AM Principles of Vibration Analysis 380 soft ware packages can be used to compute the modal vectors and natural frequencies effi ciently. For example, the solution of the eigenvalue problem of Eq. 10.5-2 can be obtained with a simple command in Matlab. Th at is, [V,D]=eig(K,M); where the diagonal elements of D are 1 2 6 2, , , and each column of matrix V contains the corresponding eigenvectors (modal vectors or mode shapes). Example 10.5-1 An engine is supported by three engine mounts on a vehicle, as shown in Fig. 10.5-1. Th e coordinates of the mass center of the engine in the global coordinate system are 802.4 mm, 5.5 mm, 616.5 mm. Th e mass and the principal moments of inertia of the engine are listed in Table 10.5-1. Th e direction cosines between the principal axes of inertia and the global axes X, Y, Z are listed in Table 10.5-2. Th e orientation of each engine mount can be defi ned by a 1-2-3 rotation sequence. Th e rotation angles and stiff ness of each engine mount are listed in Table 10.5-3. Th e coordinates of each engine mount are listed in Table 10"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001549_978-3-642-01213-6_24-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001549_978-3-642-01213-6_24-Figure2-1.png",
+        "caption": "Fig. 2 The parameters of q are defined as follows: l is the distance between the tip and the intersection point of the endoscope axis and the pivot plane, (a, b) is a displacement vector within the pivot plane, \u03b1 is the rotation angle around the z-axis (yaw), \u03b2 is the rotation angle around the y-axis (pitch).",
+        "texts": [
+            " Let us preliminary assume these openings to be fixed and so small that the endoscope can only be moved in such a way that the symmetry axis of the endoscope always passes through the center of the opening. In this case the 5d configuration space is further reduced to 3d (one translation and two rotations). The center of the opening is usually called pivot point, entry point or invariant point; we will use the first term in the following. In RAFESS we have to consider a pivot region not only a pivot point. This motivates the following definition of the configuration of the endoscope also shown in Fig. 2: q = (l, a, b, \u03b1, \u03b2 ) (4) This definition has the advantage that the user can constrain the parameters in an intuitive way. For example setting (a, b) to (0, 0) results in the 3d configuration space mentioned above. We have tracked the endoscope poses in real FESS surgeries. Based on this, we could derive the following kinematic constraints: l < 80mm, |a| < 8mm, |b| < 4mm, |\u03b1| < 45\u25e6, |\u03b2 | < 30\u25e6, (5) Environmental constraints are given by checking the configurations for possible collisions. For this purpose we need a geometrical model of the endoscope"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000612_iet-cta:20060265-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000612_iet-cta:20060265-Figure7-1.png",
+        "caption": "Fig. 7 Flexible joint manipulator configuration",
+        "texts": [
+            " It is clear from the results that the FFLC designed in this example is robust to plant parameter perturbations restricted within permissible bounds. From now on, the validity and effectiveness of the proposed controller are examined through the comparative control for a flexible joint manipulator. We examine the effects of parametric variation on behaviours of the closed-loop systems with the proposed control scheme. In order to apply the suggested controller, we need a T-S fuzzy model representation of the manipulator. Consider the single link flexible joint manipulator shown in Fig. 7, whose dynamics can be written as _x1 \u00bc x2 _x2 \u00bc MgL I sin x1 k I (x1 x3) _x3 \u00bc x4 _x4 \u00bc K J (x1 x3)\u00fe 1 J u (4:11) Fig. 6 Control results of robust FFLC (Case2) 1249 where I \u00bc 1 kgm2, J \u00bc 1 kgm2 are, respectively, the link and the rotor inertia moments, M \u00bc 1 kg is the link mass, k \u00bc 1 N/mis the joint elastic constant,L \u00bc 1 m is the distance from the axis of the rotation to the link center of mass and g \u00bc 9.8 m/s2 is the gravitational acceleration, respectively. we first transform the nonlinear system to the normal form [14] with z1 \u00bc x1 as _z1 \u00bc z2 _z2 \u00bc z3 _z3 \u00bc z4 _z4 \u00bc a(z)\u00fe b(z)u (4:12) where a(z) \u00bc \u00f049=5\u00de sin (z1)(z2 \u00fe \u00f04=5\u00de cos (z1)\u00fe 1)\u00fe ( z3 \u00f049=5\u00de sin (z1))(2\u00fe \u00f049=5\u00de cos (z1)) and b(z) \u00bc 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003473_978-3-642-20760-0_2-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003473_978-3-642-20760-0_2-Figure11-1.png",
+        "caption": "Fig. 11 Example benchmark scenario of \u2018gait learning\u2019 in modular robotics. In both configurations, an AHHS evolved that moved the robot efficiently within several tens generations.",
+        "texts": [
+            " The leading design concept of AHHS is to generate smooth fitness landscapes, that is, there is a high causality of the mutation operator (small changes in the controller result in small changes of the behavior). Several effects, for example the trade-off between evolvability and an increase of the search space, are investigated currently. First studies of the AHHS in the context of multi-modular robotics and comparisons to other controller approaches have been made [10, 29]. One of the benchmarks was the so-called \u2018gait learning\u2019 in modular robotics (see Fig. 11). One of the results, that was reported in [10], is shown in Fig. 12. It shows a comparison of the best fitness obtained by artificial evolution for N = 12 independent runs per controller approach. The superiority of AHHS over a simple artificial neural network approach is significant. The controller described above was analyzed concerning its evolvability and adaptability. As a first benchmark test a scenario was chosen in which a maze had to be explored: A robot controlled by the AHHS was put in a simulated 2D-arena and evolutionary runs were performed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000931_bfb0119383-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000931_bfb0119383-Figure5-1.png",
+        "caption": "Figure 5. The CAD model of the paper-lifting robot Fiat used for fabrication.",
+        "texts": [],
+        "surrounding_texts": [
+            "along a trapezoidal envelope. 2. r o t a t e ( f l o a t ang l e ) turns the steering wheel perpendicular to the rear wheels and rotates the robot a n g l e degrees around the midpoint of the rear wheels, the location of the sticky foot when retracted. 3. p a p e r P i c k u p ( ) winds the sticky tape spool a little, then rota tes the cam out to pick up a sheet of paper. Each rear wheel should be approximate ly at the edge of the paper. 4. p a p e r D r o p ( ) rotates the cam inward, detaching the paper from the foot. Using these primitives, the robot can move a sheet of paper from one side of a desk to the other. Since the leading edge of the paper is at least a centimeter above the lowest point on the rear wheels, the mot ion of the paper is over other sheets of paper tha t are resting on the desk. 3.2. E x p e r i m e n t s We have run two sets of performance tests on the Fia t robot. A first set of experiments was designed to quantify the reliability of shifting a single piece of paper on a desk top. Placement can be implemented as a sequential execution of p a p e r P i c k u p ( ) , r o t a t e ( f l o a t a n g l e ) , and p a p e r D r o p ( ) . A single sheet of paper was placed in front of the robot . The robot was commanded to pick up the paper, ro ta te it 180 degrees, and place it down. The location of the paper was then compared with the expected location to est imate the placement error. This experiment was repeated 54 times. We observed a 93% success rate for this cycle. Each exper iment resulted in a reasonable location for the p lacement of the paper. The 7% error were due to problems in the cam alignment for the paper lifting mechanism. We believe tha t this problem can be solved by using a be t te r shaft encoder. To measure the placement errors, we measured the location of one point (in t e rms of its x and y coordinates) and the orientat ion of the paper. We found tha t the standard deviations in the (x, y) location of the point was (0.47cm., 0.59cm.), and the s tandard deviation for the orientat ion was 3.29 degrees. Since the s tandard deviation numbers are low, we conclude tha t the experiment has a high degree of repeatibility. We also conducted an exper iment to test how well the robot can shift piles of paper. The basic experimental setup consists of a pile of three sheets of paper placed at a designated location on a desktop. The robot s tar ts at a known location and follows a pre-coded t ra jec tory to the location of the paper stack. The robot then uses the sticky foot to lift the top page and t ranspor t it to another pre-specified location on the desktop, following a given trajectory. After depositing the paper, the robot re turns to its s tar t ing location and the process is ready to be repeated. We have repeated this loop many times. We observed failures in the sys tem due to odometry. Because the robot uses no sensors to navigate, posit ion errors accumulate and result in a misalignment between the robot and the paper stack. We plan to enhance the architecture of the robots with sensors tha t will provide guidance and paper detection capabilities. 4. R e l a t e d W o r k This section reviews some previous work and its relation to the present paper. Several preceding systems have explored the connection between manipulation and locomotion. One of the earliest influential robots, Shakey [11], was a mobile manipulator. A direct approach is to attach a manipulator to a mobile platform. The JPL Cart [17] provides an early example, while Romeo and Juliet [8] provide a current example. These projects have demonstrated effective coordination of wheels and arm joints in manipulation tasks. One goal of the work described here is to explore the relation of manipulation and locomotion, a goal shared with the distributed manipulation work of Donald et al [4]. This work included a set of mobile robots pushing objects, as if each robot were a finger in a multi-fingered grasp. The OSU Hexapod [16] used dynamic stability analysis and algorithms quite similar to those sometimes used to coordinate or analyze dexterous manipulation with multi-fingered hands. The Platonic Beast [12] is probably closest in spirit to the present work. It had several limbs, each of which could be used for locomotion or manipulation. The present work can also be viewed in relation to other manipulation systems. In fact it fits naturally with work on nonprehensile manipulation. A few examples are manipulation of objects in a tilting tray [6] or on a vibrating plate [14], manipulation of parts on a conveyor belt with a single-joint robot [1], manipulation of planar objects dynamically with simple robots [10], use of a passive joint on the end of a manufacturing arm to reorient parts [13], control of an object by rolling it between two plates [3], and manipulation of planar shapes using two palms [5]. These examples are simpler than a conventional general purpose manipulator, they can manipulate a variety of parts without grasping, and they exploit elements of the task mechanics to achieve goals. In the case of the present work, each robot uses four motors to manage six freedoms (mobipulator) or seven freedoms (Fiat). Manipulation is accomplished by friction, gravity, dynamics, and adhesion, without grasping. Perhaps the most relevant previous work is the business card manipulation work of Kao and Cutkosky [7], which addressed manipulation of laminar objects by fingers pressing down from above. We still have a great deal of work to do on analysis, planning, and control, which will depend heavily on well-established techniques for non-holonomic robots. [2, 9] 5. D i s c u s s i o n a n d C o n c l u s i o n Our goal is to develop reliable desktop robots, capable of performing useful and interesting tasks on a desktop, while attached as peripherals to a desktop computer. In this paper we explore two designs for robot systems that allow the robots to move paper on a desktop in the x, y, and z directions. Both robots are small in scale and architecturally minimalist. More generally, our robots explore the deep connection between locomotion and manipulation. While several authors have noted the existence of this connection (see w the present work seeks to take the connection even further: the robot is just one of several movable objects in the task. The job of each actuator is resolved according to the task, be it manipulation, locomotion, or something not clearly classifiable as either. Our experiments show that these robots are capable of manipulating individual pieces of paper in the x, y, and z directions on a desktop 1. Many components of these demonstrations were hard wired. The purpose of the experiments was to demonstrate the basic capabilities of the robots to manipulate paper and other common objects on the desktop. We are currently incorporating a planning module in the system. Our systems are very preliminary. In the future, we plan to enhance these robots by combining their two functionalities into one mobile robot and to augment their mechanical designs with sensing. Another goal is to develop automated planners able to compute complex motions for desktop tidying tasks. We plan to incorporate (1) an external vision system to guide the robot; (2) reflective optical sensors for detecting edges of contrasting color in order to help the robot align to the edge of a sheet of paper on a contrasting background; (3) an inkjet head to be used to print barcodes on new papers arriving on the desk; and (4) optical sensors that , when combined with distance information from the shaft encoder, could be used to read barcodes on pages as the robot passes over them."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001245_j.ast.2007.05.005-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001245_j.ast.2007.05.005-Figure1-1.png",
+        "caption": "Fig. 1. Coordinated system definitions.",
+        "texts": [
+            " All analyzed results will be verified by 5-DOF simulations under large system variations and uncertainties. In following sections, effects of aerodynamic couplings are discussed and classified firstly; and then a feed forward\u2013 feedback decoupling method is proposed and evaluated from a simplified coupled system; and finally proposed decoupling method is applied to a supersonic missile flight control system with large angle of attacks. The translational and rotational dynamics of the missile shown in Fig. 1 are described by following six nonlinear differential equations [19]: U\u0307 = \u2212 q\u0304 s\u0304 m Cx \u2212 WQ + V R + Fxg m (1) V\u0307 = \u2212 q\u0304 s\u0304 m Cy \u2212 UR + WP + Fyg m (2) W\u0307 = \u2212 q\u0304 s\u0304 m Cz \u2212 V P + UQ + Fzg m (3) P\u0307 = \u2212 1 Clq\u0304s\u0304l (4) Ix Q\u0307 = Cmq\u0304s\u0304l\u0304 \u2212 Ix \u2212 Iz Iy PR (5) R\u0307 = Cnq\u0304s\u0304l \u2212 Iy \u2212 Ix Iy PQ (6) In above equations, U , V , and W are velocity components measured on the missile body axes; P , Q, and R are the components of the body angular rate; Fxg , Fyg , Fzg are the gravitational forces acting along the body axes; and Ix , Iy , Iz are the moments of inertia"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003888_s00021-012-0105-2-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003888_s00021-012-0105-2-Figure3-1.png",
+        "caption": "Fig. 3. Proof of Theorem 2.2",
+        "texts": [
+            "4), in the sense that it is constant on any solution curve. Further, we have shown in [5] that Q is positive or negative according as the net force acting on two parallel partially dipped plates as described above, and subject to the contact angles arising from the indicated solution, is attracting or repelling. This observation takes on special interest in the context of the following remark: Theorem 2.2. The horizontal force exerted on a plate by an \u201couter\u201d solution, extending to infinity, is independent of contact angle. Proof. Figure 3 illustrates a plate dipped partially into an infinite fluid at reference level 0. The lower horizontal line is the fluid at rest level, which remains unaffected by the dipping when the contact angle is \u03c0/2. We compare that configuration with that which occurs for a general contact angle \u03b3, as in the figure. Since S0 has zero curvature, we conclude from (2.2) that no net pressure is exerted on that surface. There may however be equal surface tensions \u03c3 acting in opposite directions at the end points of the segment 0P as indicated"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002352_1.3650828-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002352_1.3650828-Figure1-1.png",
+        "caption": "Fig. 1 Journal and bearing shown wi th grossly exaggerated clearance",
+        "texts": [],
+        "surrounding_texts": [
+            "L. G A L L E T T I - M A N A C O R D A\nG. C A P R I Z Centro Studi Calcolatrici Elettroniche del C. N. R. presso I'Universita Di Pisa, Pisa, Italy\ni f Misalignment in Start Lubricated Bearings The torque due to the lubricant, a nd acting on a journal in a short cylindrical bearing, is derived in the special, and particularly simple, case of a journal whirling and rocking in the neighborhood of its steady-state position under load. The results of the investigation have a bearing in the study of the damping of forced vibrations and of the development of self-excited oscillations of rotors.\nit introduction\n[ CYLINDRICAL bearing is considered (radius R, width b, radial clearance c), which is referred to a system of cylindrical coordinates (\u00a32, p, v, f ) with the f-axis along the axis of symmetry and \u00a32 in the central plane. The anomaly v is counted in that plane from an axis directed as the load acting on the journal; a first, fixed frame T is thus defined.\nA second, moving frame T\\0, x, it, z) is thought as attached to the journal, with the z-axis as axis of symmetry of the journal and 0 in the central plane of the journal. Hence, the position of the journal is completely known once the eccentricity e = j001, the anomaly (3 of \u00a320, and the Euler angles 6, ip, <p of T' with respect to T are assigned.\nThen, if rj is the constant lubricant viscosity, w the steady rotational speed of the journal in radians per unit time, and h the thickness of the film\n/,(\u201e, \u00a3) = c - e cos (u - fi) + sin (v - \\p), (1)\nReynolds' equation for the pressure p within the lubricant in a short bearing can be written as follows:\nb_\nH\nh3 bp\nL ^ \" a f .\np\nthe solution of (2), (3) is negative, phenomena of cavitation may be expected in the lubricant and the pressure must be assumed to be, approximately, zero.\nFor instance, in the steady-state case [i.e., c, /3, constant, 0 = 0], we have\n= \u20142c cos (v - /3) + (w - 2/3)e sin (v - 0)\n+ 2 f0 sin (v - + ( \u00ab - 2\\p)$d cos (v - ip). (2)\nThis equation determines a unique function p when the boundary conditions\n( V .4 sin a (1 + A cos a)3'\n(4)\n(3)\nare also required.\nThough p [as a solution of (2), (3)] is thus defined\nover all the field 0 < v < 2K, \u2014 \u2014 < f < \u2014, it has physical\nsignificance only where it is positive. Within the region where Contributed by the Lubrication Division and presented at the Lubrication Symposium, New York , N. Y . , June 6 -9 , 1965, of THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS. Manuscript received at A S M E Headquarters, February 2S, 1965. Paper No . 65\u2014LubS-9.\n[if cA and (3 are the constant values of e and fl, respectively, and a = v \u2014 f) + 7r] and the region of cavitation must be taken to be b b t h e w h o l e r e c t a n g l e 7r < a < 27r, \u2014 \u2014 < f < \u2014.\nA general solution of (2), (3) is rather complex and the problem of determining the resultant forces and moments acting on the journal and clue to the pressure field is even more difficult.\nThe task becomes instead simple if the solution of (2), (3) is sought only under conditions for the journal which correspond to small perturbations from the steady state.\nThose conditions are of special interest in the studies of shaft vibrations mentioned in the summary.\nLet us assume, then, that\ne = c{A + a{), = P + ft,\nN o m e n c l a t u r e \u2014 - \u2014 - - \u2014 a = eccentricity ratio a = e/c\nA = value of a in the steady state ai = perturbation of a from A : a \u2014 A + (Zi\nb = width of bearing c = radial clearance of bearing C\\, C3 = moments of inertia of rigid journal\ne = eccentricity of journal center: e = |Q0| F = force due to the lubricant\nand acting on the journal\nFa = components of F along the vector QO and in direction normal to \u00a320 on the journal's central plane\nh = film thickness hs = value of h in the steady state M = moment vector of the\ncouple due to the lubricant and acting on the journal\nil/e, M n = components of M along the vector QO and in a direction normal to AO on the journal's central plane\nMi, jVo = components of M along the nodal line I and in a direction normal to I on the journal's central plane\n0 = journal's center (Continued on next page)\nJournal of Basic Engineering D E C E M B E R 1 9 6 5 / 8 4 7\nCopyright \u00a9 1965 by ASME\nDownloaded From: http://fluidsengineering.asmedigitalcollection.asme.org/ on 05/12/2015 Terms of Use: http://asme.org/terms",
+            "= -\\-2ccii cos a\nd(p - ps)~ b f |_6?/co b f _\n+ 2 f t 'Ac sin a \u2014 cai sin a \u2014 2'0' sin (a + P \u2014 \\f>)\n- ( 1 - 2 i c o s ( a + /3 - i /0, ( 5 )\nwhere a prime denotes a derivative toward r. At the same time, the expression of h can be split into the sum of the steady-state value hs and a small addenda of the first order\nJ6/2 f i , m - 6 / 2 J o\nrb/2 f - Mn= f d f\nJ - 6 / 2 J 0\np sin a Relet,\np cos a Rda. ( 9 )\nFormulas (8) lead to the specification of Fe, F\u201e already given by Holmes [l ]1\nh = hs + [cai cos a \u2014 sin (a + /3 \u2014 \\p)]. (6)\nSolution of Reynolds Equation By integration of equation (5), with the specification (6) for h\nand the usual boundary conditions for f = \u00b1\u2014, one obtains for p\nthe expression\nV = + J)CO\n\u2014 R 4 c2( l + A cos a) 3\n+ (1 - cos (a + /3\n+ 6 dA s in a s i n (a -f- (3 \u2014 \\p)\n1 + A cos a: 6ai' cos a \u2014 3(2f t ' / l - ] ai) sin a\n9Aai sin a cos a 1 + A cos a\n\u00ab i ( f ) = A 2 a / - - [2d' sin (0 - 3 c\n+ (1 - 2 ^ ) 9 cos (/3 - </<)]\nF \u201e = /\n6/2 p : dt\n- 6 / 2 J O\n/ b/2 f - dt - 6 / 2 J o\n\u2022p cos a 7?(/a,\np sin a Rda\nF =\nF \u201e =\ncA2 ( i - Ar( 1\nriRb3co j 7r A\n7T 1 + 2/1 2\n2 ( 1 ^ 7 7 , \" ' '\n, 2/1(1 + A 2 ) + (1 - A) 3 *\n2A\n(10)\nc2 1 4 (1 - A 2 ) 3 / ' ( 1 - 2 f t ' ) + (1 - A 2 ) 2\n^ 2 0 ' s i n ( a + (3 - i /0\n- +)\n7T 1 + 2A2\n+ T ( i - a 2 ) ^ a i\n( 7 ) This is so because, within the limits of small perturbations, the rocking of the journal does not influence the resultant force. Change in this force from the steady-state value is related only to the parameters which describe the \"parallel\" component of journal motion.\nRocking of the journal gives rise instead to a couple, which is conversely not influenced by parallel whirl\nM. =\nUnder the circumstances envisaged here, the region filled by the lubricant (i.e., the region where (7) assigns a positive value to p) is specified by\n\u00ab l ( f ) < \u00ab \u2022 < \u00ab l ( f ) + * where\nrjRb'u ( 120c3 \\ it 1 + 5A2 ~2 (1 - A2)'/'-\n4A + (1 \u2014 A2 )2 J\n6A + r\u2014L ( i -\n(1 _ a2)3/-\nsin (j3 - + )\n4A\nOf\nA2)2 (1 - A2 )2\n7T\n6\n+\nM n = Note, in fact, that p [a i ( f ) , f ] and ; j [ \u00ab i ( f ) + ir, f ] are zero to within quantities of the second order.\nNaturally, to calculate resultant force and moment of the couple acting on the journal because of the pressure field, we need only integrate over( 0, ir) with respect to cc, the integrals over (0,0:0 and (7T,7r + ai) are quantities of the second order. The components of the force F along the vector QO, and in a direction normal to it, are given then by\n- /\n\u2022qRb'o) 120 c3\n+\n( 1 - A 2 ) 3 / *\n2/1(3 + 5/12) . 4/1 (1 - A2 )3\n1 + 2/1 2 \u00a3 + J\n\"tt 1 - 4,12\n_2 (1 - A2)3'2\n(1 - A 2 ) 2\ns in ( P - f )\n1 + 2/1 2\nV - A2)5/: 4/1\ncos (j3 \u2014 \\p) (11)\n(1 - A2 )2\nH'\ncos ( / 3 - t f O\n(8) Notice that, even during a conical whirl of small amplitude, the precession angle ip and its derivative need not., of course, be small.\n1 Numbers in brackets designate References at end of paper.\n-Nomenclature^ p = pressure within the lubri-\ncant p, = steady-state value of p R = radius of bearing\nT'(0,x,y,z) = moving Cartesian frame attached to the journal T(Q,p,p, f ) = fixed cylindrical frame attached to the bearing\nt = time a = angle counted on bearing's\ncentral plane from the direction pointing to\nmaximum film thickness in the steady state = function which defines boundary of lubricant film\n(3 = value of v on the vector Q.0\n/S = value of f3 in the steady state\n/3i = perturbation of 6 from |3: 0 = 0 + ft\nrj = constant viscosity of lubricant\n6, \\p, (p = Euler angles of T' with reference to T\nfi = center of bearing \u00ab = rotational speed of journal in radians per unit time r = nondimensional time: r =\ncof\nNote: A dot indicates a time derivative; a prime indicates a derivative toward r.\n8 4 8 / D E C E M B E R 1 9 6 5 Transact ions of the AS M E\nDownloaded From: http://fluidsengineering.asmedigitalcollection.asme.org/ on 05/12/2015 Terms of Use: http://asme.org/terms"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0003014_iros.2011.6094585-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003014_iros.2011.6094585-Figure3-1.png",
+        "caption": "Fig. 3. Robot schematic diagram. Dotted lines indicate wireless communication links.",
+        "texts": [
+            " The proposed odometry method with slip compensation is presented in Section V, and experimental validation of the odometry method is summarized in Section VI. A 978-1-61284-456-5/11/$26.00 \u00a92011 IEEE 4091 A. System Overview The omnidirectional mobile robot (Fig. 2) consists of a main body and four ASOC modules. Each ASOC module is connected to the body via a parallel linkage with shock absorbers. The robot\u2019s maximum dimension is 112 x 112 x 39 cm and the weight is 35 kg. The maximum cruising velocity is 2.2 m/s (8.0 km/h). Fig. 3 shows a schematic diagram of the robot. The onboard computer (Gumstix Overo Earth) is mounted on the main body, working as a central computing device. The onboard computer supervises all ASOC modules via Xbee wireless links, such that the onboard computer kinematically coordinates ASOC modules to move the robot in a desired direction. GPS data is collected as ground truth data for outdoor experiments. These data are sent to an operator via IEEE 802.11g, along with ASOC motion data (i.e. potentiometer for pivot angle and wheel tachometer data)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001515_sutc.2008.44-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001515_sutc.2008.44-Figure2-1.png",
+        "caption": "Figure 2. Different neighborhood types in event-based sensor networks. The outermost rectangular area represents the whole sensor field.",
+        "texts": [
+            " Since a node does not know whether it is required to be active in order to maintain good connectivity in the network, to be conservative the threshold value tends to be high to keep a large number of active nodes. That means nodes are active even when they could be asleep, and results in unnecessary energy consumption. In many event-based sensor network applications, such as habitat monitoring or intrusion detection systems, source nodes tend to locate in the same area where events occur. As shown in Fig. 2, we can categorize the sensor field into three different types: hot area, midstream area, and silent area. A node\u2019s neighborhood type is identified according to the following rules: \u2022 A node is in a hot area if it is a source or at least one of its neighbors is a source. A node is referred to as a source if it detects events and generates data packets. \u2022 A node is in a midstream area if none of its neighbors is a source and at least one of its neighbors or itself is a forwarding node. A forwarding node doesn\u2019t generate data packets but forwards packets along the routing path"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003045_s11044-012-9326-7-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003045_s11044-012-9326-7-Figure3-1.png",
+        "caption": "Fig. 3 Essential part of the telescopic-legged rimless wheel",
+        "texts": [
+            " 4 concludes this paper and describes the future research directions. 2.1 Modeling of telescopic-legged rimless wheel Figure 2 shows the model of a planar telescopic-legged rimless wheel (TRW). We call the leg frame on the floor \u201cstance leg,\u201d and assume that only the stance leg is actuated and other legs are kept at l0 [m]. We also assume that this model has a point mass, M [kg], concentrated on the central position, and the inertia moment about the CoM can be neglected. In this paper, we set \u03b1 = \u03c0/4 [rad], i.e. eight-legged, and l0 = 1.0 [m]. Figure 3 shows the essential part of the TRW which determines the stance-phase motion. This is just a variable-length pendulum and its dynamic equation becomes very simple. Let q = [\u03b8 l ]T be the generalized coordinate vector; the dynamic equation of the TRW then becomes where M(q) = [ Ml2 0 0 M ] , h(q, q\u0307) = [ 2Mll\u0307\u03b8\u0307 \u2212 Mlg sin \u03b8 \u2212Ml\u03b8\u03072 + Mg cos \u03b8 ] , S = [ 0 1 ] u, and u [N] is the control force of the linear actuator to extend/retract the telescopic-leg frame. Figure 4 shows the configuration of impact phase"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001933_s11071-010-9720-8-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001933_s11071-010-9720-8-Figure1-1.png",
+        "caption": "Fig. 1 Schematic illustration of hybrid squeeze-film damper mounted gear-bearing system",
+        "texts": [
+            "tw B bearing parameter = 6\u03bcR2L2 m\u03b43\u03c9n c radial clearance, c = R \u2212 r C damping coefficient of the gear mesh d viscous damping of disk D d m\u03c9n e static transmission error and varies as a function of time fx,fy components of the fluid film force in horizontal and vertical coordinates Fr,Ft components of the fluid film force in radial and tangential directions h oil-film thickness, h = \u03b4(1 + \u03b5 cos \u03b8) k stiffness of the retaining springs K stiffness coefficient of the gear mesh kd proportional gain of PD controller kp derivative gain of PD controller L bearing length m masses lumped at the mid-point mp mass of the pinion mg mass of the gear Og center of gravity of the gear Op center of gravity of the pinion Ob,Oj geometric center of the bearing and journal p(\u03b8) pressure distribution in the fluid film ps pressure of supplying oil pc,i pressure in the static pressure chamber Qin,i the volumetric flow rate into oil chamber i (i = 1\u20134) from the controllable orifice R inner radius of the bearing housing r radius of the journal r, t radial and tangent coordinates s speed parameter = \u03c9 \u03c9n ; \u03c9n = \u221a k m U \u03c1 \u03b4 x, y, z horizontal, vertical and axial coordinates x0, y0 damper static displacements xj , yj Xj/c, Yj/c, j = p,g \u03c1 mass eccentricity \u03c6 rotational angle (\u03c6 = \u03c9t) \u03c9 rotating speed of the shaft \u03d5b angle displacement of line ObOj from the x-coordinate (see Fig. 1) \u03a9 \u03d5\u0307b \u03b4 radial clearance = R \u2212 r , \u03b8 the angular position along the oil film from line ObOj (see Fig. 1) \u03b81 the rotating angle of gear \u03b82 the rotating angle of pinion \u03bc oil dynamic viscosity \u03b50 \u221a X2 0 + Y 2 0 \u03b2 distribution angle of static pressure region (\u00b7), (\u2032) derivatives with respect to t and \u03c6 HSFD (Hybrid Squeeze-Film Damper) has been well developed and applied in many fields. Squeeze-film damper bearings are actually a special type of journal bearing with its journal mechanically prevented from rotating but free to vibrate within the clearance space. The squeezing action produces hydrodynamic forces in the fluid film",
+            " As a result, non-linear analysis methods are generally preferred within engineering and academic circles. The current study performs a non-linear analysis of the dynamic behavior of a gear pair system equipped with hybrid squeeze-film damper bearing under gear meshing force effect. The non-dimensional equation of the gear-bearing system is then solved using the fourth order Runge-Kutta method. The non-periodic behavior of this system is characterized using phase diagrams, power spectra, Poincar\u00e9 maps, bifurcation diagrams, Lyapunov exponents and the fractal dimension of the system. Figure 1 shows a rigid gear pair supported on a hybrid squeeze-film damper (HSFD) in parallel with retaining springs. The bearing consists of four hydrostatic chambers and four hydrodynamic regions. The oilfilm supporting force is dependent on the integrated action of hydrodynamic pressure and hydrostatic pressure of HSFD. Figure 2 represents the cross section of HSFD. The structure of this kind bearing should be popularized to consist of 2N (N = 2,3,4, . . .) hydrostatic chambers and 2N hydrodynamic regions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000305_t-pas.1975.31993-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000305_t-pas.1975.31993-Figure6-1.png",
+        "caption": "Fig. 6 - Toque Profile of PM Motor with Single P k e Energization",
+        "texts": [
+            " This toque profile is validonly when the energizing current i s small such that where T, and AT are neglected. Compare (18) with (16). Equation (18) results if we replace F and Oe in (16) by NI, $ NI and 0, + e, - 45'. Thus the excitation of two phases with NI results in the shift in the detent positioa and the increase in holding toque by a factor of @for the 1 harmonic term and by a factor of 2 for the second harmonic term. Conversely, we can obtain the same holding toque by exciting one phase with fief the rated current. For high current, the toque profiles become those shown in Fig. 6. A few interesting observations can be made from Fig. 6. For K, less than 0.5, the stable and unstable points are the same as those for K, = 0. For Ka greater than 0.5, the stable points shift from the previous ones and approach 90'and 2700 as K increases, The greater the value of Ka, the greater the maxfmum toque. The maximum toque points shift 135' and 225' as K, increases. A practical implication of this phenomenon i s that with large NI a higher holding toque can be o b tained at the expense of positional accuracy. In order to increase the toque output, i t i s common prae tice to energize two sets of coils",
+            " It appears that this conclusion is valid only for a four phase stator and five tooth rotor. I believe that the order of permeance harmonic which gives rise to cogging in PM motors is given in general term by m = abs [K/N -Kl . It is also noted that this higher order permeance term has been negc ted for the computation of inductances. The results of equation 16 and Figs. 5 and 6 of the paper, although referred to, as for single phase energization, have in fact two phases, 1 and 3, of a 4-phase motor energized. The interpretation of the results of Fig. 6 presents a problem because for Ka = 0, unless PmFm is infinite, there should be no torque. Further the base torque for each curve is different. The equation [ 181 has been obtained by replacing NI byJi-NI. This method seems unjustified considering that the m.m.f. of each phase is a square or at best trapezoidal space function. In equation 16 the first term is on account of permanent magnet excitation whereas the second term is entirely on account of variation in reluctance. The reluctance torque from this equation is a second harmonic component whereas in equation 19 it is shown to be a fundamental component",
+            " Therefore, the paper neglects the higher order components in discussing specific cases. The 4th harmonic component was intentionally included to show that the cogging torque for 4-phase PM motors, which are widely used in industry, is due to this component. For other phases, the cogging torque will be due to other component as the discussor pointed out. By single phase energization, the author means that only one winding is energized. In practical motors, Coil 1 and 3 in Fig. 3 are connected in series into a one continuous winding to an appropriate power supply. Ka = 0 in Fig. 6 approximately holds when the air gap flux due to the permanent magnet dominates over that due to the energization flux. Therefore, in practical operations where the energization flux is high a pure sinusoidal torque is not attainable. The idealized mathematical model. developed in the paper gives basic insights into the operation of step motors. Next areas of investigation should include effects of eddy current and non-linear magnetic circuit. Manuscript received April 2, 1975. 151 7"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000555_485009-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000555_485009-Figure2-1.png",
+        "caption": "Figure 2. Schematic drawing of the configuration of the liquid bridge confined between a conical gripper and a flat substrate.",
+        "texts": [
+            " After this, the resemblance and differences between a cone\u2013flat and a flat\u2013flat configuration will be discussed. Based on the above results, the principle of using electrowetting for the manipulation of a liquid droplet will be presented. The experimental results as obtained from different probes will be compared with the principles. A liquid of known volume on the order of 1 nl, confined between two solids, is expected to have a uniform mean curvature since the gravitational force is negligible. If it is axially symmetric as shown in figure 2, its mean curvature \u03ba\u0304 can be described by Orr et al (1975), Lambert and Delchambre (2005b) 2\u03ba\u0304 = \u2212 r \u2032\u2032 (1 + r \u20322)3/2 + 1 r(1 + r \u20322)1/2 (1) where r = r(z) defines the shape of the liquid bridge with the following boundary conditions: r \u2032 1 = tan(\u2212\u03c0/2 + \u03b81) and r \u2032 2 = tan(\u2212\u03c0/2 \u2212 \u03b8\u2217 2 ). \u03b81 is the contact angle between the droplet and the substrate and \u03b8\u2217 2 = \u03b82 +\u03c0/2 \u2212\u03d5 and \u03b82 are the effective and true contact angles between the droplet and the gripper. The equation can only be solved numerically either indirectly by solving an elliptic integral or directly by the trial and error method (Orr et al 1975, Fortes 1982, Lambert and Delchambre 2005b)",
+            " The rupture profiles are plotted in figure 4 for various pairs of contact angles. Also given in figure 4 are profiles obtained using the arc approximation for comparison. The arc profiles are calculated by solving the following equations. The arc is conveniently described using the parameters z0, r0, R, and \u03b1, z = z0 + R cos \u03b1 r = r0 + R sin \u03b1 (3) where z0 and r0 represent the coordinates of the circular center of the arc, R is the circular radius of the arc, and \u03b1 is the angle with respect to the z axis clockwise (figure 2). \u03b11 and \u03b12 can be directly determined from the contact angles of water and the results are summarized in table 1. At the moment of rupture, only one point on the arc will contact the z axis. Depending on the contact angles, the point would be the top, bottom or middle point of the arc. The results are summarized in table 2. The volume of the liquid bridge as calculated by rotating the arc around the z axis is obtained by combining equations (2) and (3) as V = F(\u03b12) \u2212 F(\u03b11) (4) where F(\u03b1) = \u03c0 Rr 2 0 cos \u03b1 \u2212 \u03c0 R2r0(\u03b1 \u2212 sin \u03b1 cos \u03b1) + \u03c0 R3 3 (sin2 \u03b1 cos \u03b1 + 2 cos \u03b1)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003859_gt2013-95074-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003859_gt2013-95074-Figure7-1.png",
+        "caption": "Figure 7: Parameters of the linear perturbation",
+        "texts": [
+            " y y y H h = = (3) The film pressure, the film thickness and its time derivative can be written as the linear combination of a steady state part and a perturbed part. The latter one does not have a stationary part: ( , ) ( , ) sin cos , ( , ) sin cos , ( , ) ( , ) . stat stat stat stat stat stat H z H z X Y H z X Y z z X Y X Y X Y X Y \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03d5 \u03c4 \u03d5 \u03d5 = \u2212 \u2212 \u2202 \u2032 \u2032= \u2212 \u2212 \u2202 \u2202\u03a0 \u2202\u03a0   \u03a0 = \u03a0 + +   \u2202 \u2202    \u2202\u03a0 \u2202\u03a0   \u2032 \u2032+ +   \u2032 \u2032\u2202 \u2202     (4) The dimensionful parameters of the linear perturbation are illustrated in Figure 7. Substituting (4) into the generalized Reynolds equation (2) and neglecting all terms of second or higher order provides: ( ) 3 3 2 2 12 12 3 sin1 sin 2 12 3 sin , 12 stat stat stat statp x p z stat stat C p x stat stat p z H H K X z K z X H f K H z K z \u03d5 \u03b7 \u03d5 \u03b7 \u03d5\u03c1 \u03d5 \u03d5 \u03d5 \u03b7 \u03d5 \u03d5 \u03b7 \u2217 \u2217 \u2217 \u2217 \u2217    \u2202 \u2202 \u2202\u03a0 \u2202 \u2202 \u2202\u03a0   +      \u2202 \u2202 \u2202 \u2202 \u2202 \u2202           \u2202\u03a0\u2202 \u2202= \u2212 +   \u2202 \u2202 \u2202    \u2202\u03a0\u2202+   \u2202 \u2202     (5) ( ) 3 3 2 2 12 12 3 cos1 cos 2 12 3 cos , 12 stat stat stat statp x p z stat stat C p x stat stat p z H H K Y z K z Y H f K H z K z \u03d5 \u03b7 \u03d5 \u03b7 \u03d5\u03c1 \u03d5 \u03d5 \u03d5 \u03b7 \u03d5 \u03d5 \u03b7 \u2217 \u2217 \u2217 \u2217 \u2217    \u2202 \u2202 \u2202\u03a0 \u2202 \u2202 \u2202\u03a0   +      \u2202 \u2202 \u2202 \u2202 \u2202 \u2202           \u2202\u03a0\u2202 \u2202= \u2212 +   \u2202 \u2202 \u2202    \u2202\u03a0\u2202+   \u2202 \u2202     (6) 3 3 12 12 sin , stat stat stat statp x p z H H K X z K z X\u03d5 \u03b7 \u03d5 \u03b7 \u03c1 \u03d5 \u2217 \u2217 \u2217    \u2202 \u2202 \u2202\u03a0 \u2202 \u2202 \u2202\u03a0   + =      \u2032 \u2032\u2202 \u2202 \u2202 \u2202 \u2202 \u2202          = \u2212  (7) 3 3 12 12 cos stat stat stat statp x p z H H K Y z K z Y\u03d5 \u03b7 \u03d5 \u03b7 \u03c1 \u03d5 \u2217 \u2217 \u2217    \u2202 \u2202 \u2202\u03a0 \u2202 \u2202 \u2202\u03a0   + =      \u2032 \u2032\u2202 \u2202 \u2202 \u2202 \u2202 \u2202          = \u2212  (8) For the solution of these equations (5)-(8) all implicit dependencies of the other parameters and boundary conditions on the pressure have to be linearized as well, so that the application of an iterative numerical solution procedure becomes necessary"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000801_idc.2007.374545-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000801_idc.2007.374545-Figure3-1.png",
+        "caption": "Fig. 3: An example of how \u03b8i, \u03b8j , \u03b8m ij , \u03b8\u00b1m ij , \u0394\u03b8i(\u0394t) and \u0394\u03b8j(\u0394t) are calculated.",
+        "texts": [],
+        "surrounding_texts": [
+            "While the Nearest Neighbours algorithm is simple and quick to run it is not optimal in the sense that it does not locate all the emitters in the fastest possible time. The core of the Geometry Teaming algorithm capitalises on the fact that, in general, angular separation of the UAVs is more important than range from the emitter. TDOA achieves the best localisation results when the three UAVs in the team are separated in angle by 120 o around the emitter. In the following sections an outline of the cost function Cm ijk is given, which assigns costs to all UAV team and emitter pairs. Next two search techniques are detailed, describing methods of searching through the solutions generated in the cost function to find an efficient solution. Finally results of simulation are presented and the mean, standard deviation, skewness and kurtosis for histograms of the total simulation time ts are given."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001982_s12206-009-0346-z-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001982_s12206-009-0346-z-Figure1-1.png",
+        "caption": "Fig. 1. the multi-attribute simulation of windturbines.",
+        "texts": [
+            " In order to avoid this wear, gearbox manufacturers can optimize the gear geometry (lead crown and involute crown) based on virtual simulation results. Using LMS Virtual.Lab Motion, wind turbine engineers can accurately model contacts between gearbox components. This allows them to efficiently predict the transmission of loads between components like gears, shafts or bearings, while taking the flexibility of these components into account. Based on the wind load cases and using the same simulation environment, users can perform durability, vibration and noise analyses in a straightforward way (Fig. 1). In addition, the seamless integration and the preserved associativity within LMS Virtual.Lab avoid timeconsuming and error-prone transfer of data and allow development team to perform fast and multi-target optimization loops. In a standard 1.5MW wind turbine the huge input torque coming from the blades has to be transferred to realize a gear ratio between input and output shaft of more than 100 in order to match the needed rotational speed to generate electricity from the generator. Typically this is done through a 3-stage gearbox design (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001502_cdc.2008.4739357-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001502_cdc.2008.4739357-Figure8-1.png",
+        "caption": "Fig. 8. The geometric interpretation behind the algorithm proposed to visit a set of n (4 in the figure) ordered points.",
+        "texts": [],
+        "surrounding_texts": [
+            "As well as the 2 points case, to visit a certain number n of ordered point q1, q2, ...qn starting from an initial point p0 is not an easy problem. Let us denote with di,i+1 the Euclidean distance between the point i and the next desired point i + 1, where for sake of simplicity d0,1 denotes the distance between the initial position of the carrier p0 and q1. An upper bound to the optimal solution can be given for the case when the carried vehicle never takes-off, except to reach the last point, and the carrier itself visits all the points. Such an upper bound has the following total cost tn,up = [ n\u2211 i=1 di\u22121,i Vc ] \u2212 (Vc + Vv) 2Vc a\u0304 + (Vc + Vv) 2Vv a\u0304 (19) The lower bound can be reached in the same way of the 2 points case by supposing that the carrier-carried couple can be seen as a vehicle able to go at velocity Vv for a time a\u0304 for each point to reach. This would mean that the carrier is always able to \u201ccut the edge\u201d: just like in the 2 points case this optimal solution is practically reachable only for small angles between the lines connecting 2 consecutive points). Such a lower bound has the following cost tn,low = n\u22121\u2211 i=1 [ di\u22121,i \u2212 a\u0304Vv Vc + a\u0304 ] + + [ dn\u22121,n Vc + ( (Vc + Vv) 2 a\u0304 ) ( 1 Vv \u2212 1 Vc )] = n\u22121\u2211 i=1 [ di\u22121,i Vc ] + (n \u2212 1) a\u0304 ( 1 \u2212 Vv Vc ) + + ( (Vc + Vv) 2 a\u0304 ) ( 1 Vv \u2212 1 Vc ) (20) While the problem itself can be hard to solve, it is possible to build heuristic solutions able to guarantee some performances. Here we will introduce and discuss one of them based on the 1-point fastest approach with re-entry mission solution. Further possible better heuristic will be only mentioned in this paper and can be build based on the results proposed for the 2 points case. The basic idea of the \u201c1-step heuristic\u201d is the one of using, at each time step, the 1-step solution: the carrier will proceed straight to the next objective point and the vehicle will take-off only when at a distance dto = (Vc + Vv)a/2 from the objective. The carrier will keep going straight while the carried vehicle will visit the point and come back. The rendezvous point will be on the same line at distance dl = (Vv \u2212 Vc) a/2 from the objective. From this rendezvous point the ship will go towards to the next point to visit. By exploiting Carnot theorem, the total cost for such a strategy can be expressed in closed form t\u2032n = d1 \u2212 dl + n\u22121\u2211 i=2 [\u221a d2 l + d2 i \u2212 2didl cos (\u03b8i) \u2212 dl ] Vc + + \u221a d2 l + d2 n \u2212 2dndl cos (\u03b8i) \u2212 dto Vc + dto Vv (21) Where \u03b8i is the angle resulting from the line between pi\u22121pi and pipi+1. Simple computations show that the cost (21) has the interesting property of satisfying the following inequality tn,low \u2264 t\u2032n \u2264 tn,up . Inspired by the the results for the 2 points case, we propose also a second euristic solution which is based on the idea of satisfying the condition of Proposition 1 by suitably \u201dslow down\u201d the velocity of the vehicle such that the rendezvous is possible. Let us denote with \u03b8list = {\u03b81, \u03b82, ...\u03b8n\u22121} the set of the n \u2212 1 angles \u2208 [0, \u03c0] between the straight lines that connect consecutive points. For each element i of \u03b8list we compute the maximum rendezvous velocity v\u0304i according to the following algorithm \u2022 if \u03b8i \u2261 0, v\u0304i = Vv \u2022 else v\u0304i = min   Vv, \u221a 2V 2 c 1 \u2212 cos \u03b8i    The maximum rendezvous velocity v\u0304i is then used in order to bound the velocity of the vehicle to reach the desired point i following the shortest path depicted in Fig. 9. The cost of this solution can be computed as t\u201dn = n\u22121\u2211 i=1 [ di\u22121,i \u2212 a\u0304v\u0304i Vc + a\u0304 ] + + [ dn\u22121,n Vc + ( (Vc + Vv) 2 a\u0304 ) ( 1 Vv \u2212 1 Vc )] (22) Simple computations shows that the cost (22) verify tn,low \u2264 t\u201dn \u2264 tn,up . and moreover if v\u0304i = Vv for all i = 1, ..n\u22121, which in turns reduces to conditions on the angles and the ratio between the velocity of the carrier and the vehicle, then it coincides with the lower bound (20)."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001321_j.mechmachtheory.2007.08.001-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001321_j.mechmachtheory.2007.08.001-Figure4-1.png",
+        "caption": "Fig. 4. One zero degree-of-freedom planar parallel mechanism.",
+        "texts": [
+            " J OAB _QOAB \u00bc J O0A0B0 _QO0A0B0 \u00bc J O00A00B00 _QO00A00B00 \u00f021\u00de From the second row of Eq. (21), the following equation can be obtained: \u00f0 _h1 \u00fe _h2 \u00fe _h4\u00de : \u00f0 _h10 \u00fe _h20 \u00fe _h40 \u00de : \u00f0 _h100 \u00fe _h200 \u00fe _h400 \u00de \u00bc 2 : 1 : 1 \u00f022\u00de From the first row of Eq. (21), the following equation can be obtained: _h10 \u00fe _h20 \u00fe _h40 \u00bc _h100 \u00fe _h200 \u00fe _h400 \u00bc 0 \u00f023\u00de The above two equations mean that three degrees-of-freedom must be all translational. One zero degree-of-freedom planar parallel mechanism is showed in Fig. 4. If the coordinates of points O, A, B, C, D and E are ( 0.25, 0.15,0), ( 0.25,0.4, 0), ( 0.1, 0.25,0), ( 0.1, 0.05, 0), (0.05, 0.2,0) and (0.15, 0.35, 0) in turn and the reference point is A, then singular value decomposition of coefficient matrix is: V \u00bc 2:56 0 0 0 0 0 0 1:53 0 0 0 0 0 0 0:62 0 0 0 0 0 0 0:33 0 0 0 0 0 0 0:27 0 0 0 0 0 0 0:01 2 666666664 3 777777775 \u00f024\u00de The Eq. (24) means that rank of coefficient matrix is full, i.e., the degree-of-freedom is zero. If the chain OA has three links like Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001419_iciea.2009.5138489-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001419_iciea.2009.5138489-Figure5-1.png",
+        "caption": "Fig. 5 The master manipulator structure",
+        "texts": [
+            " The robot can work in pure master-slave operation or master-slave and local-automatic operation. In master-slave operation mode, the system is controlled by an operator who moves a master manipulator, which has six degrees of freedom., The movements are transmitted to a slave manipulator, and force signals - produced when the slave interacts with the environment are transmitted back to the operator so he has tactile sense of the conditions under which the slave is working. The master manipulator structure is shown in Fig. 5. The system is composed of a dual-arm slave manipulator system ,which has 6 degree of freedom in each slave arm and a 6-axis force/torque sensor is mounted at the tip of each arm.. In master-slave and local-automatic operation mode, the live working of robot process follows special steps. The dualarmed manipulator is lifted to a corresponding height under master-slave operation mode. Next, after Changing operation mode from the masterservo to the local-automatic operation, and choosing a certain kind of live workings pre-defined, the dual-armed manipulators can do the kind of live working under the trajectory program adjusted according to the trajectory program predefined and position information from sonar"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001199_0278364907085565-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001199_0278364907085565-Figure11-1.png",
+        "caption": "Fig. 11. Tree-structured modular manipulator.",
+        "texts": [
+            " The results obtained in both cases are generally comparable, because the achieved motion performances are sensibly determined by the overall number pn of processing phases thus, an over-employment of a small set of modules is generally equivalent to a reduced employment of a large set of modules. In this section the extension of the previously devised control architecture and algorithms to the more general case of a treestructured kinematic chain is provided. For ease of exposition, but without any loss of generality, reference will be made to the structure sketched in Figure 11, obtained by joining three serial chains, named \u201ca\u201d, \u201cb\u201d and \u201cc\u201d, analogous to the chain seen in the previous section. As the composing robotic modules (identical to those developed at Genoa Robotic And Automation Laboratory (GRAAL) and described in the next section) are characterized by single-face docking connectors, a special Y-shaped module (hereinafter called a \u201cY-module\u201d) has been inserted at the junction point of the three chains in order to enable a branched connection to be established",
+            "comDownloaded from As concerns its functional role, the Y-module (or any other module occasionally becoming the joining element of a branched chain) has to perform some additional computations (other than those required by the procedure described in previous sections) for dealing with the presence of the branching point. Nevertheless, it will be shown that even such additional processing burden is quite limited. Therefore, its implementation may be easily provided onboard all of the PCUs (because every module might at a certain time \u201cbecome\u201d a Y-module) without a significant effect on the architectural considerations drawn at the end of previous section. To better describe the additional functionalities that must be provided by the Y-module, consider again the structure in Figure 11, whose task clearly consists of making tool frames at and bt (associated with the upper sub-chains) asymptotically converge toward goal frames ag and bg , respectively7. To accomplish such a task in a distributed fashion, analogously to what happens in the case of serial chains, at every sampling interval a preliminary data-exchange phase has to be performed among all of the PCUs transmit information of a geometric nature. In this case, however, the presence of the Y-module introduces two slight modifications on the procedure seen in previous sections: (i) during the forward phase the Y-module now has to evaluate two \u201cinternal\u201d transformation matrices in order to compute matrices ya T and yb T which are then sent in parallel to the first PCUs of sub-chains \u201ca\u201d and \u201cb\u201d, respectively (ii) during the backward phase three parallel data-exchange processes are independently performed by the PCUs of the three sub-chains more specifically modules of sub-chains \u201ca\u201d and \u201cb\u201d transmit the position of their corresponding tool frame at and bt , respectively modules of sub-chain \u201cc\u201d instead exchange just the position of the tool frame of their sub-chain, ct , which coincides with the base frame of Y-module",
+            " the other modules, despite the structural redundancy of the manipulator, are not able to compensate for the absence of joint 6 in this case. In this way the trajectory followed by the end-effector becomes non-linear. Otherwise, if the posture allowed the other modules to produce a contribution compensating for the temporary inoperativeness of joint 6, the resulting trajectory would be sufficiently linear (as compatible with 50 iterations). As concerns tree-structured chains, many simulations have been executed by considering different models of modular manipulators similar to that depicted in Figure 11. In particular, hereafter the results obtained on a model of a structure with 20 degrees of freedom are reported. More specifically, the considered robot is characterized by two upper sub-chains with eight degrees of freedom mounted on a common lower sub-chain with four degrees of freedom. In order to underline the behavior of the proposed technique once applied to tree-structured chains, a sufficiently high number of iterations, namely 100, have been used. In the first simulation (see Extension 3) the pair of endeffectors is asked to move from points [0"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000296_acc.2006.1657581-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000296_acc.2006.1657581-Figure5-1.png",
+        "caption": "Fig. 5. Behavior of the non-actuated link (q2, q\u03072) generated by twisting algorithm.",
+        "texts": [
+            "1 [rad] for the joints 1 and 2, respectively. For the Pendubot, we select A1 = 0.15 [rad] as amplitude of oscillations and \u03c9 = 20 [rad/s] (see Bode diagram in Fig. 3 for the open-loop system). For the twisting algorithm (13) we derive the following parameters: c1 = 50, c2 = 5.0. (31) Fig. 4(a) shows simulation results for the output variable. Fig. 4(b) gives a better picture of the amplitude and frequency of the oscillations by plotting the steady-state output trajectory between 8 and 10 secs. In the phase portrait of Fig. 5 we show the periodic motion of the non-actuated link. In Fig. 6 we illustrate the LPRS behavior for the Pendubot using twisting algorithm. Frequency of periodic oscillations found as solution of (23) is \u03a9 = 20.38 [rad/s]. Bode Diagrams This paper has addressed the problem of orbital stabilization of a underactuated system around an unstable equilibrium point. The generation of periodic motion is made through the second-order sliding mode (SOSM) twisting algorithm. In this case the transient process may converges to a periodic motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000836_6.2008-7302-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000836_6.2008-7302-Figure1-1.png",
+        "caption": "Figure 1. Planar engagement geometry.",
+        "texts": [
+            " The 3D motion of such a missile can be separated to two perpendicular planes, which can be solved independently. The solution to one such plane will be presented in this paper. The 3D motion is derived by combining the solutions in both planes. We first present the full non-linear kinematics equations of the interception problem, which will be used in the non-linear simulation in Sec. VI. Then, the linearized equations which will be used to derive the proposed guidance law will be presented. In Fig. 1 a schematic view of the planar endgame geometry is shown, where XI \u2212 OI \u2212 ZI is a Cartesian inertial reference frame. We denote the missile and target by the subscripts M and T , respectively. The speed, normal acceleration, and flight path angles are denoted by V , a, and \u03b3, respectively; the range between the adversaries is r, and \u03bb is the angle between the LOS and the XI axis. The remaining terms will be defined in the next subsection. Neglecting the gravitational force, the engagement kinematics, expressed in a polar coordinate system (r, \u03bb) attached to the missile, is r\u0307 = Vr (1a) \u03bb\u0307 = V\u03bb/r (1b) where the speed Vr is Vr = \u2212 [VM cos (\u03b3M \u2212 \u03bb) + VT cos (\u03b3T + \u03bb)] (2) and the speed perpendicular to the LOS is V\u03bb = \u2212VM sin (\u03b3M \u2212 \u03bb) + VT sin (\u03b3T + \u03bb) (3) During the endgame, the target and missile are assumed to move at a constant speed",
+            " We assume that the dynamics during the endgame can be represented by arbitrary order linear equations x\u0307M = AMxM + BMuM (5a) \u03b3\u0307M = aM/VM (5b) where aM = CMxM + dMuM (6) and xM is the state vector of the interceptor\u2019s internal state variables with dim(xM ) = n. We denote the part of the acceleration without dynamics, if it exists, as the direct lift, while the part with dynamics as the specific force. These equations can represent the closed loop missile dynamics. They can also represent the linearized open loop dynamics in which case the obtained guidance law will actually be an integrated guidance-autopilot controller. The endgame kinematics used for the guidance law derivation was presented in Fig. 1. The X-axis, aligned with the LOS used for linearization, is denoted as LOS0. z is the relative displacement between the target and the missile normal to this direction. The target and missile accelerations normal to LOS0 are denoted by aTN and aMN , respectively; and the intercept angle is (\u03b3T +\u03b3M ). We denote the required value of (\u03b3T +\u03b3M ) as xc 4. The derivation of the guidance law in this paper will be performed based on a linearized model. If during the endgame the missile and target deviations from the collision triangle are small, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002698_s1068798x10040234-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002698_s1068798x10040234-Figure1-1.png",
+        "caption": "Fig. 1. Simulation of the adaptive technological control system in ensuring specified surface roughness Ra: RC, rated characteristic Ra = f(s); s is the supply.",
+        "texts": [
+            " (2) is based on the mathematical model [3] (3) Ra C0sx v y ,= Ra_izm Ra_zad \u0394Ra_zad+\u2265 Ra_izm Ra_zad \u0394Ra_zad\u2013\u2264 v const= Cv Ra_izm sx = s Ra_izm( ) Ra_zad Cv \u239d \u23a0 \u239b \u239e 1/x = \u23ad \u23aa \u23aa \u23aa \u23aa \u23ac \u23aa \u23aa \u23aa \u23aa \u23ab , Ra_izm Ra_zad \u0394Ra_zad+\u2265 Ra_izm Ra_zad \u0394Ra_zad\u2013\u2264 s const= Cs Ra_izm/v y= v Ra_izm( ) Ra_zad/Cs( )1/y= \u23ad \u23aa \u23aa \u23aa \u23ac \u23aa \u23aa \u23aa \u23ab , Ra k0 s k1 90 \u03b3+( ) k4 r k2 v k3 .= We reduce Eq. (3) to the form in Eq. (1), by adopt ing constant values of r and (90 + \u03b3) and setting v = 180 m/min. Then (4) where k0 = 70, k1 = 0.85, k2 = 0.37, k3 = 0.39, and k4 = 0.13 are coefficients of the model; r = 0.8 mm is the radius at the cutter tip; \u03b3 = \u201310\u00b0 is the front angle of the cutter [3]. Suppose that we require Ra_zad = 3.2 \u03bcm, with a tolerance \u0394 = \u00b110%. To this end, we must establish sup ply s = 0.133 mm/turn, which corresponds to point 1 on the rated characteristic (Fig. 1). As a result of the external perturbations, the surface roughness exceeds the tolerances. The measured value Ra_izm may be either larger than (Ra_zad + \u0394Ra_zad) or less than (Ra_zad \u2013 \u0394Ra_zad); this corresponds, respectively, to points 2 and 3 in Fig. 1. In other words, the rated characteristic may be shifted upward or downward under the action of the external perturba tions and will pass either through point 2 or through point 3 (Fig. 1). The displacement of the rated charac teristic may be due to change in the coefficients of the model under the action of the external perturbations. In simulation, we derive three characteristics. The first is the initial or rated characteristic. In construct ing the second and third characteristics, the coeffi cients in their mathematical models will change pro portionally relative to the rated model. For the second characteristic, all the coefficients are less than the rated values; for the third characteristic, they are higher than the rated values. Consider the case where Ra_izm = 3.707 \u03bcm, which is greater than (Ra_zad + \u0394Ra_zad) (point 2 in Ra k0s k1 v k\u2013 3; k0 k0 90 \u03b3+( ) k4 r k2 ,= = 420 RUSSIAN ENGINEERING RESEARCH Vol. 30 No. 4 2010 SUSLOV, PETRESHIN Fig. 1). We determine the new (corrected) supply to ensure the specified roughness on the basis of Eq. (2) Thus, to ensure the specified roughness Ra_zad = 3.2 \u03bcm (in machining conditions that are varying randomly), we need to reduce the supply to s = 0.112 mm/turn (point 2 in Fig. 1). When Ra_izm = 2.749 \u03bcm, which is less than (Ra_zad \u2013 \u0394Ra_zad) (point 3 in Fig. 1), we find that Thus, to ensure the specified roughness Ra_zad = 3.2 \u03bcm (in machining conditions that are varying ran domly), we need to increase the supply to s = 0.159 mm/turn (point 3 ' in Fig. 1). The temperature and force fields in the machining zone have considerable influence on the characteris tics of the surface layer [4, 5]. Their instability during machining (on account of fluctuation of the margin, the hardness of the blank, the tool wear, and other fac tors) results in constant variation in the machining force and temperature. Therefore, a stochastic model may be adopted to relate the surface properties and the state variables Pz and T of the machining process FMP = where FMP is the physicomechanical parameter of the surface layer; Pz is the primary component of the cutting force, N; T is the temperature in the cutting zone, \u00b0C; Cf, xf, yf are coefficients of the model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003045_s11044-012-9326-7-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003045_s11044-012-9326-7-Figure11-1.png",
+        "caption": "Fig. 11 Four holonomic constraint forces",
+        "texts": [
+            " Summarizing the above seven velocity constraint conditions of (42), (43), (46)\u2013(48), the matrix J I (q) \u2208 R 7\u00d78 is specified as follows: J I (q) = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 1 0 J13 J14 \u22121 0 J17 J18 0 1 J23 J24 0 \u22121 J27 J28 0 0 0 0 1 0 0 0 0 0 0 0 0 1 0 0 0 0 0 1 0 0 0 0 0 0 0 0 0 0 0 1 0 0 1 0 0 0 \u22121 0 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 , (49) where J13 = l1 cos \u03b81, J14 = sin \u03b81, J17 = \u2212l2 cos \u03b82, J18 = \u2212 sin \u03b82, J23 = \u2212l1 sin \u03b81, J24 = cos \u03b81, J27 = l2 sin \u03b82, J28 = \u2212 cos \u03b82. We can finally accomplish the transition by replacing q\u0307+ 1 with q\u0307+ 2 in q\u0307+ = [ q\u0307+ 1 q\u0307+ 2 ]T. Matrix J (q) in Eq. (35) is also derived as J (q) = \u23a1 \u23a2\u23a2\u23a3 1 0 J13 J14 \u22121 0 J17 J18 0 1 J23 J24 0 \u22121 J27 J28 1 0 0 0 0 0 0 0 0 1 0 0 0 0 0 0 \u23a4 \u23a5\u23a5\u23a6 =: \u23a1 \u23a2\u23a2\u23a3 J 1(q) J 2(q) J 3 J 4 \u23a4 \u23a5\u23a5\u23a6 . (50) The holonomic constraint force vector in Eq. (35) can be divided into J (q)T\u03bb = J 1(q)T\u03bb1 + J 2(q)T\u03bb2 + J T 3 \u03bb3 + J T 4 \u03bb4, (51) and each undetermined multiplier \u03bbi corresponds to the forces in Fig. 11. The \u03bb4 is identical to the vertical ground reaction force, and X-position of the ZMP is then calculated by Xzmp = \u2212uA \u03bb4 . (52) 3.2 Output following control We choose the two telescopic-leg lengths and the relative hip angle as the control output. The output vector, y \u2208 R 3, is then defined as y = \u23a1 \u23a3 b1 b2 \u03b8H \u23a4 \u23a6 = STq, (53) and its second-order derivative with respect to time becomes y\u0308 = STq\u0308 = A(q)u \u2212 B(q, q\u0307), (54) where A(q) := STM(q)\u22121Y (q)S, (55) B(q, q\u0307) := STM(q)\u22121 ( Y (q)h(q, q\u0307) + J (q)TX(q)\u22121J\u0307 (q, q\u0307)q\u0307 ) "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001199_0278364907085565-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001199_0278364907085565-Figure5-1.png",
+        "caption": "Fig. 5. Velocity contribution and velocity error vectors computed at the second iteration by (a) VPU 1 and (b) VPU 2.",
+        "texts": [
+            " Although q2 1 induces a tool velocity term, x2 1, contributing to the execution of the original velocity task, part of , corresponding to the error vector 2 1 1 1 x2 1 I h2h# 2 1 1 (16) still remains unaccomplished. Therefore, with the aim of fulfilling such a residual part, a second iteration, identical to the first that has just concluded, can be started by asking VPU 1 to provide a tool velocity contribution maximally similar to vector 2 1. To this aim, VPU 1 (via the application of (11) to input 2 1) finds a second joint velocity term q1 2, which provides an additional tool velocity contribution, x1 2, and leads to an error vector 1 2 (see Figure 5(a)). This is therefore sent as a reference vector to VPU 2, which in turn finds q2 2, inducing x2 2 and leaving the unaccomplished part 2 2 (see Figure 5(b)). A number of other iterations (say K) can then be performed in the same manner, before the end of the overall external sampling interval is reached. At that point every VPU evaluates the total joint velocity reference by summing up the set of terms computed during all of the iterations namely qi K k 1 qi k i 1 2 (17) The resulting joint velocity reference qi is then sent to the underlying low-level joint controller, and the overall procedure can be started again, working on a new tool velocity reference vector "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001371_cts.2008.4543971-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001371_cts.2008.4543971-Figure1-1.png",
+        "caption": "Figure 1. One of the UAV Platforms Used to Fly Search Missions Guided by the Proposed Search Algorithm.",
+        "texts": [
+            " In this paper we consider multiple UAVs persistently searching a specified area and focus on a decentralized search algorithm capable of achieving superior search performance through collaboration between team members. The proposed search algorithm, originally introduced in [1], allows each UAV to determine its own heading based on a weighted combination of four independent factors: (1) distance from neighboring UAVs, (2) distance from approaching search boundaries, (3) elapsed time since a subsection of the mission area was last visited, and (4) current heading. Although the algorithm itself has already been demonstrated in real-world tests involving groups of two UAVs such as the one shown in Figure 1 [9], the independent contribution of each of the four aforementioned factors in the heading calculation has not yet been the subject of accurate analysis in the context of the joint search efficiency. This paper\u2019s contribution lies on an in-depth statistical analysis of the impact of each factor that affects the computation of future UAV heading directions in the proposed search al- 978-1-4244-2249-4/08/$25.00 \u00a92008 IEEE 506 gorithm and how those factors affect the overall search effectiveness. To measure the efficiency of a combined search effort by multiple UAVs, three metrics are employed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003352_9781118067147-Figure6.2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003352_9781118067147-Figure6.2-1.png",
+        "caption": "Figure 6.2 (a) Commercial SPR system with desktop instrument, microfluidics, and biochip (Biacore, GE Healthcare). (b) Another commercial SPR system with integrated prism and chip (SPREETA, Texas Instruments). Reproduced with permissions from [4] and [10].",
+        "texts": [
+            " This basic principle has allowed the construction of compact systems with disposable biochips, for the user either to graft with their own choice of antibodies (or any other interacting biomolecules for a particular purpose) or to buy preimmobilized chips [4]. 86 SURFACE PLASMON RESONANCE BIOSENSOR DEVICES SPR has been used by many experimenters for a variety of investigations ranging over applications such as antibody epitope mapping, kinetic evaluation, food and clinical testing of toxicants, drug interactions, and so on [8,9]. Presently, approximately 10 vendors of SPR instruments provide the market with alternative devices of varying design. Figure 6.2 shows two INTRODUCTION 87 extremes of these: the successful Biacore system from GE Healthcare [4] and the small and compact SPREETATM system developed by Texas Instruments [11]. The Biacore system is an expensive benchtop instrument with an advanced sensor chip construction, automated fluid handling in microfluidic system, a cartridge and a multipurpose software program for evaluation of signals. The SPREETA system is a compact unit with the sensor chip integrated into the prism, light source, and detector [6,12,13,14]",
+            " Alternative D places the prism function inside the flow cell. The other functions remain unchanged. More units inside the flow cell probably require space inside the compartment. But it is also a challenge for further miniaturization of the separate units. Here, need for a novel design can be felt. In alternative E, the D alternative is taken further by also placing the irradiation and detection functions inside the flow cell. This means a compact design solution. It envisages a design similar to the SPREETA system shown in Figure 6.2b. In the alternative F, the C configuration has been multiplied. This could benefit from again having the flow distribution placed separated from the flow cell, at least a part of it. The number of repeated flowcell/surface units canvary (five is depicted in the figure). For example, a Biacore 1000TM instrument has 4 channels and a Biacore A100TM has 20 channels. The six alternatives shown in the figure are a starting point. More alternatives can be generated. However, it is clear that a deeper analysis of the MECHATRONIC DESIGN APPROACH OF SPR SYSTEMS 91 configuration is required"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001890_j.electacta.2009.09.019-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001890_j.electacta.2009.09.019-Figure1-1.png",
+        "caption": "Fig. 1. Exposed view of the disposable glucose test strip. 1-polypropylene sheet; 2-working electrode (effective area: 2 mm \u00d7 2 mm); 3-reference/counter",
+        "texts": [
+            " Hydrophilic fumed silica Aerosil380, D50 = 7 nm) was purchased from Degussa (Essen, Ger- any) and Dow Corning 1510 antifoam and Dow CELLOSIZE QP 40 rom Dow Corning (Shanghai, China). Glucose stock solution, with concentration of 18,000 mg/dL in normal saline, was prepared at east 24 h before required and allowed to maturate at room temperture. Potassium ferricyanide, hydroxyethyl cellulose and all other hemicals were analytical-reagent grade and obtained from Beijing hemical Reagents Company (Beijing, China). All solutions were repared with Millipore deionized water. .2. Preparation of disposable glucose test strip The disposable glucose test strip as shown in Fig. 1 was manuactured as following. On one flat side of polypropylene sheet (1), carbon electrode array comprising a set of isolated and disconected electrode pair (2 and 3), conducting traces (4 and 5) and ontacts (6 and 7) were screen-printed using Electrodag 423SS caron ink (Acheson, Michigan, USA) and then dried at 80 \u25e6C for 60 min. hen an insulating film (8) was subsequently screen-printed using lectrodag 451 SS insulating ink (Acheson, Michigan, USA) on the onducting film. The dielectric layer (8) covered portion of working lectrode (2), reference/counter electrode (3), conducting traces (4, ), but did not cover electric contact (6, 7)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002050_iros.2010.5650727-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002050_iros.2010.5650727-Figure6-1.png",
+        "caption": "Fig. 6. Concept of predictive force feedback",
+        "texts": [
+            " Even though the transmission delay of force information is relatively small, there is still a time delay between the movement of the operator and the force presentation This time delay may result in damage to organs since information of the contact with an organ is delivered to a surgeon with this time delay. Thus, we propose that force information be presented with a prediction to overcome this time delay. To present force information without a time delay, it is necessary to predict the moment of contact with the organ. To predict contact with an organ, a virtual area is introduced around the organ to detect the approaching instrument (Fig. 6). Assuming that the position and shape of the organ are known and fixed, a virtual area is set around the organ. At the moment the surgical instrument enters the virtual area, a virtual force is presented to the operator. In this system, the positions of the instruments are estimated from kinematics and the virtual areas are defined in Cartesian coordinates. This approach makes it possible to present force information earlier than the presentation of information of actual contact and thus the surgeon knows the moment of contact with organs with a smaller time delay"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003069_transducers.2011.5969384-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003069_transducers.2011.5969384-Figure2-1.png",
+        "caption": "Fig. 2. Fabrication process. (a) Fixed electrode deposition. (b) Parylene deposition and sacrificial layer patterning. (c) Parylene deposition. (d) Moving electrode deposition and permalloy electroplating. (e) Additional parylene layer deposition and aluminum mask deposition. (f) Sacrificial layer removal and diaphragm release.",
+        "texts": [
+            " In the reference chamber, the fluid viscosity and vibration damping are glucose-independent because of a lack of glucose binding by PAM and only vary with environmental parameters (e.g., temperature). Capacitive sensors that contain fixed electrodes on the device substrate and moving electrodes integrated with the diaphragms measure the differences in diaphragm vibration between chambers, allowing glucose detection in the face of environmental fluctuations. The device fabrication process began with the deposition and patterning of chrome (5 nm) and gold (100 nm) to form fixed electrodes (500 \u00b5m \u00d7 500\u00b5m) on a thermally grown SiO2 layer on a silicon wafer (Fig. 2a), followed by the deposition of a Parylene passivation T2C.007 978-1-4577-0156-6/11/$26.00 \u00a92011 IEEE Transducers\u201911, Beijing, China, June 5-9, 2011970 layer. A sacrificial photoresist layer (5 \u00b5m) was then spin-coated and patterned to define electrode air gaps (Fig 2b), followed by the deposition of a Parylene layer (Fig. 2c). A second layer of chrome (5 nm) and gold (100 nm) were next deposited for the moving electrodes and permalloy seed layer. Subsequently, defined by a photoresist pattern (5 \u00b5m), strips of permalloy (220 \u00b5m \u00d7 35 \u00b5m \u00d7 3 \u00b5m ) were electroplated (Fig. 2d). This was followed by the removal of the photoresist, patterning of the moving electrodes (500 \u00b5m \u00d7 500\u00b5m), and deposition of an additional Parylene layer (3 \u00b5m) for passivation. Eight etching holes (250 \u00d7 250 \u00b5m) were opened through the Parylene layers by reactive ion etching to expose the sacrificial photoresist (Fig. 2e) , which was subsequently removed by acetone (80 \u00b0C) to release the diaphragms (520 \u00b5m \u00d7 520\u00b5m) (Fig. 2f). These etching holes were then sealed by epoxy. After wafer dicing (Fig. 3a and 3b) and wire bonding, a chip was bonded to an SU-8 sheet (thickness: 100 \u00b5m), in which holes of appropriate sizes were patterned to define the microchambers (1 mm in diameter) as well as the inlets and outlets for polymer solution handling. This SU-8 sheet was in turn bonded using epoxy to a regenerated cellulose acetate semi-permeable membrane with a molecular weight cutoff of 3,500 Da and thickness of 20 \u00b5m. Chemicals and reagents used in the experiments include PHEAA-ran-PAAPBA, PAM, and D-(+)-glucose (Sigma-Aldrich)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003688_s11633-010-0554-0-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003688_s11633-010-0554-0-Figure1-1.png",
+        "caption": "Fig. 1 Image of a desired trajectory composed of a sine curve",
+        "texts": [
+            " In this paper, a recurrent neural network is applied for a feedforward controller of PUMA560 manipulator, which is a representative of industrial robots with six degree-offreedom. Hereafter, the recurrent neural network is called RNN. The feedforward controller includes gravity compensator and Coriolis/centrifugal force compensator. Also, a simple and adaptive learning technique is proposed for a robotic servo system to deal with large-scale joint driving torques. The learning technique and control effectiveness are evaluated through simulations using the dynamic model of PUMA560 manipulator as shown in Fig. 1. The remainder of this paper is organized as follows: Section 2 explains the computed torque control method used as a model-based servo system. Then, teaching signals for RNNs are calculated. In Section 3, integrated RNNs are designed to compensate nonlinear terms such as gravity term and Coriolis/centrifugal force term. Furthermore, a simple and adaptive learning technique by using a scaler is proposed for large-scale teaching signals processed in back propagation algorithm, then the effectiveness is discussed",
+            " The computed torque control method is used for nonlinear control of an industrial manipulator, which is composed of a model based portion and a servo portion. The servo portion is a closed loop with respect to the position and velocity. On the other hand, the model based portion has the inertia term, gravity term and Coriolis/centrifugal force term, which work to cancel the nonlinearity of the manipulator. In order to realize high control stability, the position and velocity feedback gains used in the servo portion should be tuned suitably. In this section, a simple desired trajectory as shown in Fig. 1 is calculated for the computed torque control method. The computed torque control method generates joint driving torque to follow the trajectory. The desired trajectory and joint driving torque are used as teaching signals for neural networks. The dynamic model of a manipulator with friction term is generally given by M (\u03b8)\u03b8\u0308 + H(\u03b8, \u03b8\u0307) + G(\u03b8) + F r(\u03b8, \u03b8\u0307) = \u03c4 (1) where M (\u03b8) \u2208 R6\u00d76 is the inertia term in joint space. H(\u03b8, \u03b8\u0307) \u2208 R6\u00d71 and G(\u03b8) \u2208 R6\u00d71 are the Coriolis/centrifugal force term and gravity term, respectively",
+            " The nonlinear compensation terms M\u0302 (\u03b8), H\u0302(\u03b8, \u03b8\u0307) and G\u0302(\u03b8) are calculated by using the recursive NewtonEuler formulation[14] to cancel the nonlinearity and are effective to achieve stable trajectory control. The block diagram of the computed torque control method is shown in Fig. 2. Although the gravity term and Coriolis/centrifugal force term are ordinarily obtained by computing the inverse dynamics through recursive Newton-Euler formulation, the computational load within a sampling period is considerably large. As an example, the desired trajectory in Cartesian space has been designed as shown in Fig. 1. The tip of the robot arm draws a sine curve in x-y plane. The trajectory in Cartesian space is resolved into the trajectories in joint coordinate system by using the inverse kinematics of PUMA560, which are shown in Fig. 3. Also, Fig. 4 shows the joint driving torques needed for realizing the trajectory shown in Fig. 1, which are obtained through a trajectory following simulation by using (2). The simulation was carried out by using the dynamic model of PUMA560 manipulator on Matlab system[14\u221216]. The robotic dynamic model with the friction torque term given by (1) was applied. The friction torque term F r(\u03b8, \u03b8\u0307) consists of the viscous friction torque and Coulomb friction torque, which is represented by F r(\u03b8, \u03b8\u0307) = BG2 r\u03b8\u0307 + Gr\u03c4 c{sign(\u03b8\u0307)} (3) where B is the coefficient matrix of viscous friction at each motor, Gr is the reduction gear ratio matrix which represents the motor speed to joint speed, and \u03c4 c{sign(\u03b8\u0307)} is the Coulomb friction torque, which appeared at each motor",
+            " Therefore, the array of the networks represents a dynamic model to compensate the nonlinear terms consisting of Coriolis/centrifugal force term and gravity term, which can be seen by comparing Fig. 2 with Fig. 11. Fig. 11 shows the block diagram of the control system including the integrated RNNs, in which the torque vector \u03c4 NN from the integrated RNNs is multiplied by S\u22121 to call back the original scales of teaching signals. The control system is similarly applied to the trajectory following control problem as shown in Fig. 1. Fig. 12 (a)\u2013 (f) show the control results of six joints, respectively. It is observed that each joint can be successfully controlled along the desired trajectory although there exist small delays between the desired joint angles and controlled ones. In this case, the elements of the output \u03c4 F B from the servo system without gravity term and Coriolis/centrifugal force term are shown in Fig. 13 (a)\u2013(f), respectively. Note that \u03c4 F B is the output only from the feedback control with Kp and Kv, which does not include the compensation \u03c4 NN using the RNN"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001736_0022-2569(68)90007-4-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001736_0022-2569(68)90007-4-Figure8-1.png",
+        "caption": "Figure 8. Dynamic balancing with two identical rotating masses.",
+        "texts": [
+            " Let us introduce the following three conditions: the trajectory of the first counterweight m' t (Figs. 7 and 8) and two expressions, which guarantee the maintenance of static balancing of the mechanism : m l x l = - m l x l m l z l = - , n l z l . Satisfying the above expressions assures that the center of mass Ua of the dynamic counterweights will be stationary. If we take identical counterweights m'2=m't=m', then in the first case (Fig. 7) the trajectory of the center of the second counterweight will also be straight and parallel to the first. In the second case (Fig. 8), the trajectories will coincide. In both instances the centers of tile counterweights will lie on the opposite sides of a straight line passing through point u , . Finally, let us turn to statico-dynamic balancing, which must be brought about simultaneously. Since the number of equations to be satisfied equals three, the number of free parameters will be equal to k = 2 n ' - 3. (9) Let us examine how mechanisms are balanced statico-dynamically with different numbers of counterweights. The problem cannot be solved with one counterweight since two unknown variables cannot satisfy the conditions of three equations"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001451_978-0-387-09643-8-Figure9.13-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001451_978-0-387-09643-8-Figure9.13-1.png",
+        "caption": "Figure 9.13: (a) Visualization of a Sliced Gaussian Mixture Density Consisting of a Gaussian Mixture in xk Subspace and a Dirac Mixture in \u03b7k Subspace. (b) The efficient update and the re-approximation of the resulting Gaussian mixture density by a sliced Gaussian mixture density is alternately performed(adapted from [132]).",
+        "texts": [
+            " The resulting equations of the prediction step and measurement step in the case of the identification of model parameters (SRI method) can be found in [134]. The Sliced Gaussian Mixture Filter basically consists of three steps: the decomposition of the estimation problem, the utilization of an efficient update, and the reapproximation of the density representation. 1. Decomposition: The nonlinear high-dimensional estimation problem is decomposed into a linear high-dimensional problem (state estimation) and nonlinear low-dimensional problem (parameter estimation). This can be achieved by means of the sliced Gaussian mixture density; see Figure 9.13 (a). 2. Efficient update: Based on the decomposition in a linear and nonlinear estimation problem both the prediction step and the measurement step can be performed with an overall more efficient performance. Basically, analytic and efficient estimators, such as Kalman filter, are exploited to efficiently perform the estimation update in the linear high-dimensional subspace. The estimation in the nonlinear subspace is performed by nonlinear estimators, such as the Dirac mixture filter, visualized in Figure 9.13 (b). 3. Re-approximation: The estimation based on the sliced Gaussian mixture densities leads to a density representation consisting of Gaussian mixtures in all subspaces. In order to bound the complexity, the resulting density needs to be re-approximated by means of the sliced Gaussian mixture density, depicted in Figure 9.13 (b). The remainder of this section is devoted to a more detailed description on the three steps of the Sliced Gaussian Mixture Filter (SGMF). Furthermore, the update equations for the simultaneous reconstruction and node localization (SRL method) are given. Decomposition through sliced Gaussian mixtures For the decomposition of the nonlinear estimation problem into a (conditionally) linear and a nonlinear problem, we proposed in our previous research work [92] [134] so-called sliced Gaussian mixture densities as a density representation",
+            " This density function f(xk, \u03b7k) is represented by a Dirac mixture in the nonlinear sub-space \u03b7 k (parameter space) and Gaussian mixture (a)Sliced Gaussian mixture densities (b)Efficient update and reapproximation in the linear sub-space xk (state space), f(xk, \u03b7k) = M\u2211 i=1 \u03b1ik \u03b4(\u03b7 k \u2212 \u03bei k )\ufe38 \ufe37\ufe37 \ufe38 Dirac mixture \u00b7 f(xk|\u03be i k )\ufe38 \ufe37\ufe37 \ufe38 Gaussian mixture , (9.21) where \u03b4(\u00b7) is the Dirac delta function and \u03b1ik their respective weighting coefficients. The density parameters \u03bei k \u2208 Rs can be regarded as the position of the individual density slices, as shown in Figure 9.13 (a). The marginal density in nonlinear subspace \u03b7 k is given by a Dirac mixture function, according to f(\u03b7 k ) = M\u2211 i=1 \u03b1ik\u03b4(\u03b7 k \u2212 \u03bei k ), (9.22) where \u03b1ik and \u03bei k represent the weights and positions of the Dirac functions, respectively. The density representation along the individual slices is assumed to be a Gaussian mixture density f(xk|\u03be i k ) = Ni\u2211 j=1 \u03b2ijk N ( xk \u2212 \u00b5ij k , Cij k ) , (9.23) with \u03b2ijk , \u00b5ij k \u2208 Rr, and Cij k \u2208 Rr\u00d7r denoting the weights, means, and covariance matrices of the j-th component of the Gaussian mixture density of the i-th slice",
+            " The resulting equations of the prediction step and measurement step in the case of the identification of model parameters (SRI method) can be found in [134]. The Sliced Gaussian Mixture Filter basically consists of three steps: the decomposition of the estimation problem, the utilization of an efficient update, and the reapproximation of the density representation. 1. Decomposition: The nonlinear high-dimensional estimation problem is decomposed into a linear high-dimensional problem (state estimation) and nonlinear low-dimensional problem (parameter estimation). This can be achieved by means of the sliced Gaussian mixture density; see Figure 9.13 (a). 2. Efficient update: Based on the decomposition in a linear and nonlinear estimation problem both the prediction step and the measurement step can be performed with an overall more efficient performance. Basically, analytic and efficient estimators, such as Kalman filter, are exploited to efficiently perform the estimation update in the linear high-dimensional subspace. The estimation in the nonlinear subspace is performed by nonlinear estimators, such as the Dirac mixture filter, visualized in Figure 9.13 (b). 3. Re-approximation: The estimation based on the sliced Gaussian mixture densities leads to a density representation consisting of Gaussian mixtures in all subspaces. In order to bound the complexity, the resulting density needs to be re-approximated by means of the sliced Gaussian mixture density, depicted in Figure 9.13 (b). The remainder of this section is devoted to a more detailed description on the three steps of the Sliced Gaussian Mixture Filter (SGMF). Furthermore, the update equations for the simultaneous reconstruction and node localization (SRL method) are given. Decomposition through sliced Gaussian mixtures For the decomposition of the nonlinear estimation problem into a (conditionally) linear and a nonlinear problem, we proposed in our previous research work [92] [134] so-called sliced Gaussian mixture densities as a density representation",
+            " This density function f(xk, \u03b7k) is represented by a Dirac mixture in the nonlinear sub-space \u03b7 k (parameter space) and Gaussian mixture (a)Sliced Gaussian mixture densities (b)Efficient update and reapproximation in the linear sub-space xk (state space), f(xk, \u03b7k) = M\u2211 i=1 \u03b1ik \u03b4(\u03b7 k \u2212 \u03bei k )\ufe38 \ufe37\ufe37 \ufe38 Dirac mixture \u00b7 f(xk|\u03be i k )\ufe38 \ufe37\ufe37 \ufe38 Gaussian mixture , (9.21) where \u03b4(\u00b7) is the Dirac delta function and \u03b1ik their respective weighting coefficients. The density parameters \u03bei k \u2208 Rs can be regarded as the position of the individual density slices, as shown in Figure 9.13 (a). The marginal density in nonlinear subspace \u03b7 k is given by a Dirac mixture function, according to f(\u03b7 k ) = M\u2211 i=1 \u03b1ik\u03b4(\u03b7 k \u2212 \u03bei k ), (9.22) where \u03b1ik and \u03bei k represent the weights and positions of the Dirac functions, respectively. The density representation along the individual slices is assumed to be a Gaussian mixture density f(xk|\u03be i k ) = Ni\u2211 j=1 \u03b2ijk N ( xk \u2212 \u00b5ij k , Cij k ) , (9.23) with \u03b2ijk , \u00b5ij k \u2208 Rr, and Cij k \u2208 Rr\u00d7r denoting the weights, means, and covariance matrices of the j-th component of the Gaussian mixture density of the i-th slice"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003727_j.engfailanal.2011.06.004-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003727_j.engfailanal.2011.06.004-Figure3-1.png",
+        "caption": "Fig. 3. Typical mesh model of (a) U-bolt and (b) assembly of the blade and U-bolts.",
+        "texts": [
+            " The fracture surface of the U-bolt was investigated by visual inspection and scanning electron microscopy (SEM), using secondary electron detectors. A structural analysis of U-bolts of a cooling tower fan is performed using finite element modeling. SolidWorks software is used for geometry modeling of the cooling fan including fan blade, hub plate and U-bolts as shown in Fig. 2. Finite element software ANSYS 12 is used for stress analysis. The elements used in U-bolt model are all 8-node solid brick elements Solid45. Finite element model for stress analysis is consisted of 10,000 elements and 16,200 nodes for each U-bolt as shown in Fig. 3. All parts of the cooling fan assembly except the hub plate are considered flexible. Since blade model must transfer forces to the U-bolts, it is not considered rigid in this analysis. To model the contact between the U-bolts, blade and hub surfaces, CONTA174 and TARGE170 elements are used. In this analysis, CONTA174 elements are used to model contact surfaces for bolt\u2013blade, bolt\u2013hub plate, and blade\u2013hub plate contacts. Load applied to the finite element model is chosen according to the information of Hawden cooling fans data sheet for 9900 ENF 6 types"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000816_icma.2008.4798779-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000816_icma.2008.4798779-Figure2-1.png",
+        "caption": "Fig. 2 Concept of support system for forward tilting posture",
+        "texts": [
+            " has been developing so called \u201cSmart Suit\u201c for the same purpose though [10], assist power is conveyed in the direction from the shoulder to the hip and it is not directly transferred to supporting torque for forward tilting. As the result, only 13% support at maximum is realized. In this paper, Section 2 describes the concept in order to realize forward tilting support, and Section 3 explains configuration of the support system for real use. Section 4 illustrates estimation results about how effective the support system is by the experiment to keep forward tilting posture with load. As shown in Fig.2 (a), to lift from the back is one of convincing ways to support forward tilting posture. A wearer can put one's full weight on it in this manner. Because forward tilting motion is a kind of rotation one in terms of the upper and lower body around the waist, mechanical structure to realize it is developed as shown in Fig.2 (b). In this figure, inner black circle is fixed on the body and outer circle connected to the back frame rotates around it. The McKibben artificial muscle is chosen as an actuator for its light weight, flexibility, and large output. Because the McKibben artificial muscle shortens in straightaway, it is mounted at the back frame and by using a wire, it connects to outer circle by using a outer for a wire. Note that the outer is fixed to back frame as shown in Fig.2 (c) so that linear motion of the McKibben artificial muscle is converted to rotation motion. By controlling the elongation of the McKibben artificial muscle, forward tilting motion is supported. How to fix inner black circle on the body in Fig.2 (b) is very crucial for this mechanism, since if it rotates, support torque for forward tilting is not obtained. Fig.2 (c) describes the solution, i.e., use of waist belt and thigh belt. Waist belt is used as a normal belt for pants. When the McKibben artificial muscle is constricted, inner black circle is forced to rotate with respect to back frame. Waist frame directly connected to inner black circle is applied and by using thigh belt through connecting belt, waist frame is fixed in order to avoid rotation. In this mechanism, the front part of thigh is burdened by thigh belt during forward tilting support. Let\u2019s here investigate how much load the front part of thigh has",
+            ", t t b MgT TbMg \u03b8 \u03b8 \u03b8\u03b8 cos sin cossin \u22c5=\u2234 \u22c5=\u22c5 (3) The load for horizontal and vertical direction of the front part of thigh ( xT and yT respectively) are as follows; \u03b8\u03b8 \u03b8 \u03b8\u03b8 sinsin cos sinsin Mg cb MgTT t t tx =\u22c5== (4) \u03b8\u03b8 \u03b8 \u03b8\u03b8 sincos cos sincos Mg bb MgTT t t ty =\u22c5== (5) As a result, one leg receives 2xT and 2yT . Let\u2019s here check the concrete value. Because of sin function, 2xT and 2yT will be maximum when 2\u03c0\u03b8 = . From the average information of adult body build, M , , b , c denote as 30kg, 0.3m, 0.2m, 0.26m respectively. Thus maximum values will be ][5.22][5.2202 ][3.17][6.1692 kgfNT kgfNT y x == == . (6) These values seem to be heavy load. Whereas as we described in Fig.2 (c), thigh belt has an area, i.e., about 0.1m x 0.2m corresponding to the front part of thigh. From this point of view, loads of horizontal direction and vertical one will be 86.5gf and 112.5g per 1cm2. It seems to be not so hard. In face, weraers feel it is bearable. A. McKibben Artificial Muscle As mentioned above, the McKibben artificial muscle is applied as an actuator. The McKibben artificial muscle was developed in the 1950s and 1960s for artificial limb research [11]. It is small, lightweight, simple, soft, flexible, and has no stiction [12]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001727_0022-2569(69)90006-8-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001727_0022-2569(69)90006-8-Figure2-1.png",
+        "caption": "Figure 2. Input locking positions.",
+        "texts": [],
+        "surrounding_texts": [
+            "263\ntogether with the primary parts of\nsin ~t2 sin ~., 3 cos 02 = (cos it., cos z2_, 3 - cos z~ 3-~ cos ~ t ) + sin z23.t sin z2.~ t cos 0.~ (2.4)\nsin ~z3 sin~3, cos 03=(cos ~2~ cos ~34--cos ~4l COS 7t2) +sin ~ t s i n ~t~ cos 0t (2.5}\nIn each of the above equations the dual angles and the trigonometric functions involving them can be expanded into primary and dual components in the following form:\nO~=Oi+eSi ~'ij = ~'ij + 8aij (2.6}\nsin Oi= sin 0~+ ~Si cos O~ cos Oi= cos Oi--eSi sin O~ sin c2U = sin ~i j+gai j cos ~ij cOS ~. i j : COS ~ i j - 8aii sin ~ij (2.7)\nIn any dual number equality the primary and dual components on each side of the equality must be equal. Equations (2.1)-(2.3) can therefore be expanded into primary and dual components to yield six trigonometric equations involving the sixteen parameters of the mechanism. The six equations give the two possible configurations of the mechanism for each value of the input rotation 0~. The primary parts of equations (2.4) and (2.5), which are obtained by removing the dual symbol signs, are needed to specify each configuration uniquely.\n3. Classification of the R - C - C - C Mechanism (i) Lockhlg positions We differentiate equation (2.1) with respect to time in order to obtain a relationship\nbetween the input and output dual velocities 0~ and 04. This relationship reduces to either\na quadratic for cos 0t(b t =0) or for cos 04(0~=0), the roots of which are the cosine laws\nfor a spatial triangle [2]. By successive substittttion in the remaining displacement equations\nwe obtain three cosine laws and two sine laws for each triangle.\nFor the input link in a locking position we obtain:\ncos 01 - cos at_, cos a . , - cos(a23 + a3.0 sin at 2 sin a4t\n(3.1)\ncos 02 - cos ~2t2 cos(a23 + a~) - cos zL t sin ~t,_. sin(,2, 3 + a34)\n(3.2)\ncos 03 = + 1 (3.3)\ncos 04 - cos ~.~1 cos(c223 + a34) - cos ~1 z (3.4) + sin ~4 t sin(a23 + a34)\nsin 02 = sin d4t sin 0t (3.5) sin(a23 4- a3D",
+            "264\nsin 03 = 0 (3.6)\n\u00b1sin c21z sin 0 t MnO,- s i ~ + z 2 3 4 ) (3.7)\nand for the output link in a locking position:\nCOSOl =C0S(~12\"['~23)C0S~41 --C0S~34 (3.8) sin(~z2-t-~23)sin$,t\ncos 02= \u00b11 (3.9)\n03 -- COS(~12 + ~ , 3 ) C 0 S ~ 3 4 - - C 0 S ~ 1 C O S - - \"\n\u00b1sin(~lz\u00b1~z3)sin ~34 (,3.10)\n0 cos~34cosa4.1-c0s(02,2+~23) COS 4 --\nsin ~23. ~ sin a.~l (3.11)\nsinOz=O (3.12)\nsin 03 = +sin c241sin 0l ~3.13) sm ~34\nsin 0.~ - sin(:21 z + ~23!sin Ot (3.14) sin c23~\nIn order to determine a locking position uniquely it is necessary to satisfy the six trigonometrical equations of the cosine laws for the spatial triangle together with the primary parts of the sine laws. The dual parts of the sine and cosine laws are given in the Appendix. The primary parts are identical to the dual trigonometrical expressions above with the dual symbol removed.\nThe input and output links may lock in either of the two positions shown in Figs. 2 and 3. For an input locking position links a23 and a34 become collinear whilst for an output locking position links a~, and a, 3 are collinear.",
+            "S2 \\ 3~\nS4 / t\nF i g u r e 3. Ouput locking positions.\n265\nGenerally locking conditions will not be reached in practice. This is because an arbitrary choice of the nine fixed parameters au , :qj and S , is unlikely to satisfy simultaneously the trigonometrical equations of either the input or output dual sine and cosine laws.\n(ii) Limit positions of rotation Input and output limit or stationary values of rotations are obtained by satisfying only the primary parts of the dual cosine laws. Here the links a23, a34 and a l . , az3 become parallel with the displacements $2 and $3 having non-zero values.\nIt follows that the R - C - C - C mechanism may be classified into crank-rocker, doublerocker or drag-link using the following relationships:\nhtpttt\nCOS 01 -- COS ~t 12 COS ~4.1 - - C0S(~23 \"l- 0134 )\nsin ~, 2 sin \"s4,\n02 COS 0[ 12 COS(~2 3 -'}\" 0~34) - - COS ~41 COS - -\ns,n ~,_, sin(or 23 ---~34)\ncos 03 = __+ l\ncos 04 - cos cz 4 ~ cos(~23 +- ~t 3.,) - cos ~t -' (3.15) +s in ~ql sin(ct23 \"F ct3a.)\nOutput\ncos 01 - cos(cq 2 -+ ct23)cos ~,,t - cos ct3,,\nsin(~t t 2 + ~t23)sin ~4t\ncos 02 = +_ 1\ncos 03 -- c\u00b0s(ct t 2 q--- ct2a)cos ct3.* -- cos ct 4 + sin(~q 2 ----- ~t23)sin ct34\ncos 0.~ -- cos ct 3,, cos ot 4 t -- cos(or t 2 q\" ct23) sin ~34 sin ~t.~ t\n(3.16)"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000199_j.ijimpeng.2006.07.005-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000199_j.ijimpeng.2006.07.005-Figure1-1.png",
+        "caption": "Fig. 1. The FE model.",
+        "texts": [
+            " This is due to the nonlinear boundary condition defined between the slider and disk during the head lift process. An FEM of the ST drive from Seagate is created with the commercially available software, ANSYS 8.1. The overall model includes the pivot bearing, the HSA and the disk. The HSA is composed of the following parts: the over mold with inlaid coil, two arms and two head suspension assemblies. The two head arm assemblies are integrated together with a screw and shim. The details of the head gimbal assembly are also simulated. The overall FE model is shown in Fig. 1. The Cartesian coordinate, X, Y and Z are parallel to the directions of the drive length, width, and thickness, respectively. The over mold, pivot shaft, screw, shim, arm, slider are meshed with solid elements while the hinge, load beam, gimbal and the disk are meshed with shell elements. The pivot bearing is simulated with 36 spring elements. The whole FE model includes 9710 solid elements and 2374 shell elements. The suspensions are initially bent with a certain angle at the middle line of the hinge to produce proper preloading forces at the head\u2013disk interface"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002808_s12239-011-0037-5-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002808_s12239-011-0037-5-Figure1-1.png",
+        "caption": "Figure 1. Schematic diagram of in-situ torque and angular acceleration measurement.",
+        "texts": [
+            " In addition, the angular acceleration was measured at the same position as the torque measurement, to compare the dynamic characteristics of the angular acceleration with the torque variation. The moment of inertia of the overall powertrain system, including the engine and the transmission, and the mean torque were determined experimentally. The idle stability was evaluated through the measured torque and angular acceleration, and the results were compared. 2. INSTRUMENTATION 2.1. Torque Measurement Figure 1 shows the torque sensor installation position and experimental setup. The torque sensor is located between an engine cylinder block and torque-converter, and the engine brake torque can be measured directly without modification of the engine\u2019s mounting structure in a vehicle. The output signal from strain gauges that are attached to the torque sensor is digitized by an analog to digital converter and a microprocessor. The digitized signal is transmitted wirelessly through a Bluetooth module. A telemetry system is essential because the torque sensor rotates at a very high speed. The Bluetooth module that receives the transmitted data is fixed on the cylinder block, and finally the data are transferred to a PC via a USB interface (Lee, 2007). 2.2. Angular Acceleration Measurement Designing the circumferential part of the torque sensor as ring gear shape, allowed the angular acceleration to be measured at the same position as the torque measurement. The magnetic pickup is fixed adjacent to the torque sensor, as shown in Figure 1. The output signal from the magnetic pickup is transferred to a PC after going through a timer and an analog to digital converter, as shown in Figure 2, and the angular acceleration was calculated. In this study, the torque and the angular acceleration were measured for two different SI engines. The torque sensor was made individually for each engine and was calibrated against a standard torquemeter. Engine A is a 3.3 Valvetrain type Type 1 (direct acting) Type 2 (end pivot) Variable valve timing (Intake) Yes Yes Variable valve lift (Intake) No Yes Variable valve timing (Exhaust) Yes Yes L, V6 type, and engine B is a 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000901_978-3-540-73422-2_4-Figure4.24-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000901_978-3-540-73422-2_4-Figure4.24-1.png",
+        "caption": "Fig. 4.24. Un-reduced path executed by the Nomad XR4000. The vertex d of the new box shaped object B forces a slow down near it.",
+        "texts": [
+            " The velocity profile for the same is shown in Figure 4.22-b. Figure 4.23 is the counterpart of Figure 4.22 where the robot adapts its path to a better time-length reactively. The big dips in the velocity profile of Figure 4.22-b are filtered in Figure 4.23-b considerably as the robot avoids the obstacles with larger separation. The time reduced execution tallied to 10.9s while the un-reduced version was executed in 12.5s. The trajectory time at planning was 7.9s. The above graphs are those obtained in simulation. Figure 4.24 shows the un-reduced executed path by the XR4000 Nomadic robot in our laboratory at LAAS. The obstacles in the original map are shown by black lines, while the segments perceived by the SICK laser are shown in lighter shades of gray. Some of these segments get mapped to the ones in the map and the others are considered new segments. This is done by a segment based localization algorithm. The segments of concern here are those which form a box shaped obstacle marked B in Figure 4.24. The vertex d of this obstacle casts a shadow on the robot\u2019s sensory field, which forces it to slow down at those locations due to Equation 4.2. The execution time for this un-reduced path is 10.6s. The time reduced counterpart is shown in Figure 4.25 that tallied to 9.6s. The original planning time was 8.8s in the absence of the box shaped object. The velocity profile for the same is shown in Figure 4.26. A proactive safe planning algorithm and its reactive version that facilitates realtime execution has been presented"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000772_robio.2007.4522451-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000772_robio.2007.4522451-Figure3-1.png",
+        "caption": "Fig. 3 Fluid model of the robot",
+        "texts": [
+            " As long as the negative pressure comes into being between outside and inside of the suction cup, the thrust creates, and then the robot is pushed on the wall by the atmospheric pressure and the thrust. In the next, we mainly discuss the cause of the negative pressure in detail. The definite air gap leakage is needed, so the air gap leakage in the air-sealed ring of the cup is an important key parameter. Supposing the conditions as follows [1]: (a) There is no rotating airflow at the air-sealed surface between the sealed ring and the wall surface. (b) The airflow leakage consists with radial direction along the wall surface. (c) The Mach number Ma<0.3, namely. See Fig.3. In the suction cup, we set up cylindrical coordinates (r z). 0zv v , and gas continuity equation: 0r rv v r r (1) The parameters have no relation with , therefore, according to Navier-Stokes equation [9]: 2 2 2 1 r rv vP r z r (2) 0P P z g (3) Where, - Gas density (kg/m3). - Aerodynamic adhesion coefficient (Pa.s). Give the boundary condition: 0z , 0rv ; z h , rv U , U is the airflow velocity where is in the max gap. So it can be given: 1 2r dP Uv z z h z dr h (4) Equation (4) is composed of two parts: one is because of the different pressure; another is because of the shearing flow"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001139_1.3159368-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001139_1.3159368-Figure1-1.png",
+        "caption": "Fig. 1 Turbocharger rotor assembly with SFRB support",
+        "texts": [
+            " Predicted total shaft motion is also in good agreement with test data for all engine loads and over the operating TC shaft speed range. The comparisons validate the rotor-bearing model and will aid in reducing product development time and expenditures. DOI: 10.1115/1.3159368 Introduction Turbochargers enable smaller and more fuel-efficient passenger ehicle internal combustion IC engines with power outputs comarable to those of large displacement engines. In a TC, exhaust ases drive a turbine wheel, which is connected to a compressor heel by a thin shaft, as shown in Fig. 1. The compressor forces denser charge of air i.e., larger air mass into the engine cylin- ers, improves combustion efficiency, thus producing more power. hile simple in concept, TC design and operation are complex. s a TC rotor spins, it vibrates due to several causes imbalance, ngine-induced vibrations, fluid film bearing self-excitations, erodynamic loading, pressure fluctuations, etc. . The vibration mplitudes need to be controlled to ensure TC reliability. The bearings support system is a key feature of a reliable turocharger design since bearings greatly affect the TC static and ynamic forced performances",
+            " In addition, unstable operaion results when the external shock frequency is close to that of he natural frequency of the RBS. Lee et al. 19 also demonstrate oth analytically and experimentally that vibration amplitudes ere amplified when the external shock frequency was close to a ystem natural frequency. This paper presents progress on the refinement of the linear and onlinear rotordynamic models developed in Refs. 9\u201313 to preict the response of a TC supported on a SFRB due to specified ase excitations that simulate engine-induced vibrations. Turbocharger Structural Model and SFRB Thermal odel Figure 1 displays the TC rotor assembly with the SFRB as a ne-piece design integrating the compressor and turbine bearings t its ends. A button pin prevents ring rotation. Figure 2 shows the C rotordynamic model consisting of 43 finite elements for the otor, including the thrust collar and spacer, and 13 finite elements or the SFRB. Childs 20 presents the theoretical aspects emloyed for rotor lateral bending and associated finite element odeling. Validation of the rotordynamics model requires good orrelation between measured and predicted rotor physical proprties, as well as free-free natural frequencies and mode shapes"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001822_cphc.200800659-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001822_cphc.200800659-Figure7-1.png",
+        "caption": "Figure 7. Wire electrode assembly for CNT applications. a) Printed circuit board, b) electrode wire (25-mm gold or platinum), c) copper conductor strip, d) conducting silver paint. Insulating cover not shown.",
+        "texts": [
+            " A home-built radio-frequency heating generator was coupled to the electrochemical cell and controlled by the PC via the microcontroller interface. For electrophoretic experiments and for electric heating of the covered wires, an ordinary power supply was used, in either the voltage-stabilized or current-stabilized mode. Electrochemical investigations were carried out in a 50-mL glass cell equipped with a PTFE lid in which the reference electrode (Ag/AgCl in KCl sat.), counter electrode (a Pt wire in glass tube) and working electrode were mounted. The working electrodes (Figure 7) were 12-mm-long pieces of gold or platinum wire, 25 mm in diameter. The metallic conductors were isolated by a standard laminating foil and by means of PTFE tape. Raman spectra were excited by the 2.41-eV line of an Ar+ laser (Innova 300 series, Coherent) and recorded by a T-64000 spectrometer (Instruments, SA) interfaced to an Olympus BH2 microscope (50 LWD objective). The laser power impinging on the cell window was between 4 and 5 mW. The Raman spectrometer was calibrated before each set of measurements by using the F1g line of Si at 520"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001533_2007-01-2234-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001533_2007-01-2234-Figure4-1.png",
+        "caption": "Figure 4. Thrust bearings highlighted to show axial load path",
+        "texts": [
+            " The gear contact analysis includes a consideration of the calculated gear mesh misalignment, a non-linear tooth stiffness, as well as the detailed micro-geometry surface definition. The predicted position of the contact on the tooth flank will influence the effect of the gear mesh force. For example, if the contact is away from the centre of the tooth, this will generate a turning moment on the gear, which will have to be reacted by the supporting bearings. The rear planetary gear set in this example is shown in Figure 2. Within the model, the axial load is properly taken by the addition of rolling element thrust bearings (Figure 3), as shown in Figure 4. The clutch and brake parts within the system are modeled as mass, inertia and stiffness components. The completed model is shown in Figure 6, including all gears, bearings, planet carriers and clutches. The boundary conditions of the analysis are then defined. This entails specifying the input speed and torque (or power), and which clutches or brakes are locked. The convergence scheme is shown in Figure 7. The non-linear analysis can be completed in a couple of minutes. The results show the full six degree of freedom deflection of the system, the rotational speeds of and forces acting on all components, gear and bearing misalignments"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002357_acc.2012.6315526-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002357_acc.2012.6315526-Figure1-1.png",
+        "caption": "Fig. 1. Simplified quadrotor model representation at hovering and the coordinate systems (the body and the earth reference frames).",
+        "texts": [
+            "00 \u00a92012 AACC 2515 tion recovery control strategy is presented in Section IV. Simulation results are presented in Section V followed by concluding remarks in Section VI. The quadrotor consists of four dc motors on which propellers are mounted in a cross configuration. Each propeller is connected to the motor through reduction gears. All the propellers axes of rotation are fixed and parallel. The front and the rear propellers rotate counter-clockwise, while the left and the right ones turn clockwise. In Figure 1, the schematic of a simplified quadrotor structure is shown where \u2126i(rads\u22121) refers to the propellers rotational speed. While at hovering, all the four propellers rotate at the same speed to counterbalance the acceleration due to gravity. Basic movements are achieved by the differences between the propellers speeds. Vertical rotation is achieved by creating an angular speed difference between the two pairs of rotors. Increasing or decreasing the speed of the four propellers simultaneously permits climbing and descending"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002405_s12541-012-0292-z-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002405_s12541-012-0292-z-Figure1-1.png",
+        "caption": "Fig. 1 Dimension over balls Mdk on helical external gear teeth with an odd number of teeth",
+        "texts": [
+            " \u03b1Kt is pressure angle in transverse section at circle through center of ball, deg.. mn is normal module, mm. \u03b2 is helix angle, deg.. By comparing formula (1) and formula (4), it is found that cos(90/ z) is eliminated in formula (4). The explanation is expressed as follows. Because of the parallel measuring element faces the two measuring pins are forced on a helical path out of the position occupied by the measuring balls in the dimension over balls and into position which are exactly opposite one another over the center point of the gear tooth system, (In figure c of Fig. 1, the factor cos(\u03c0 / 2z) is eliminated, and formulae 3 applies), the dimension over pins is twice as large as the radial single-pin dimension. The helical motion of the pins is opposed so that the measuring points move apart axially as a function of the helix angle \u03b2M and the angle (\u03c0 / 2z). For this reason, the measuring element faces shall be held parallel to the gear teeth axis, as is also necessary with the dimension over balls. The gear tooth system shall have a certain width bmin, as shall the measuring elements so that the pins do not protrude beyond the tooth ends, as shown in Fig. 1. And Minimum face width = measuring width + margin, as shown in formula (5). (5) where bmin is minimum face width, mm. Pt is transverse pitch on the reference cylinder, mm. \u03b2 is helix angle, deg.. DM is diameter of ball used for measurement, mm. \u03b2M is helix angle at measuring circle, mm. For the variable \u03b2M of formula (5), it is not given in the standard. The calculation of the variable may refer to the general theory of pin and ball measurement of chapter 2 in Vogel W. F.\u2019s book, the Md mnz \u03b1tcos \u03b2 \u03b1Mtcoscos ------------------------------- 90 z -----\u239d \u23a0 \u239b \u239ecos DM+= inv \u03b1Mt inv \u03b1t DM mnz \u03b1ncos -------------------------- 2 \u03b1n\u03c7tan z --------------------- \u03c0 2z -----\u2013+ += Mdk dk DM+= d \u03b1tcos \u03b1Ktcos ----------------- Mdk mnz \u03b1tcos \u03b2 \u03b1Ktcoscos ------------------------------- DM+= bmin Pt 2 tan \u03b2 ---------------- DM sin \u03b2M DM+ += proposition III is expressed as: Contact position and contact points of a ball in a thread groove are identical with those of a measuring pin of equal diameter"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000293_tmag.2006.871428-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000293_tmag.2006.871428-Figure6-1.png",
+        "caption": "Fig. 6. Five-phase 10=8 switched reluctance motor. Phase 1 (stator poles 1 and 6) under excitation.",
+        "texts": [
+            " 3 for first-order interpolation, with two phases connected in series with a 10-A excitation, show good results for either LM and MB. But LM produces a more stable result as shown in the detail of Fig. 4. If the interpolation order is augmented to two (using hierarchic interpolation on and hierarchic complete elements on moving band), the results presented in Fig. 5 are obtained. In this case, against expectations, we obtain a worst result with more oscillations than with first-order interpolation. Because the torque calculation in thin air gaps may surprise, the switched reluctance motor [4] shown in Fig. 6, which has an air gap of 0.25 mm, is also simulated. For this machine, two discretizations (Fig. 7) are used: mesh 1 (with 2486 elements, 1605 nodes, 120 edges on and four layers of elements in the air gap) and mesh 2 (with 4861 elements, 2730 nodes, 240 edges on , and 2 layers of elements in the air gap). The air-gap thickness g is 0.25 mm. The interface is placed in the middle of the air gap. The MB is placed in the layer next to the rotor side of . The torque is always calculated in the layer of quadrilateral elements placed next to the stator side of "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003694_mawe.201100858-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003694_mawe.201100858-Figure4-1.png",
+        "caption": "Figure 4. Parameters influencing the life of components in the battery electrical vehicle.",
+        "texts": [
+            " The requirements for structural strength had been adapted to the usage behaviour for small as well as electric vehicles [1]. Figure 3 shows the classification of the single trip length for different cars: BMW 1series and MINI Cooper with internal combustion engine as well as MINI E in Berlin. It is shown, that there is only a small difference in usage behaviour [1]. During the design of components, particularly those relevant to safety, many diverse parameters influencing the service life of a battery electrical vehicle must be taken into account, Fig. 4. The most important parameters influencing the service life of a vehicle are: \u2013 The design of components in terms of their geometry and shape: radii, stiffness, resistance to bending, torsion, etc. \u2013 Surfaces formed in the production process influencing notch factors. \u2013 The materials with their specific material properties (tensile and compression strength, flexural stiffness, etc). \u2013 Loading encountered during service conditions (service, spe- cial event and misuse loads) \u2013 Usage behaviour by the customer, e"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002886_1548512911414951-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002886_1548512911414951-Figure2-1.png",
+        "caption": "Figure 2. The flight modes of Tiltrotor UAV.",
+        "texts": [
+            " These powerplants are like propellers on an airplane, which provide thrust with a tilt angle. The control variables, both rotational speeds 1, 2 RPM\u00f0 \u00de and tilt angles \u03b81, \u03b82 degrees\u00f0 \u00de of the propellers can be controlled independent of each other. Tiltrotor UAV, having the physical properties given in Table 1, is designed with the graphical user interface (GUI) programs explained in Section 7. The transitions between the flight modes are accomplished by changing the rotational speeds of the propellers and the angular positions of the rotors, as shown in Figure 2. As both of the rotors are tilted forward progressively, the orientation of the fuselage becomes parallel to the Earth\u2019s surface for level flight. The main components of Tiltrotor UAV, environmental effects and gravity are modeled as shown in Figure 3. The medium that the aircraft experiences has many variables such as temperature, pressure, density, viscosity, speed of sound, and gravitational acceleration. These variables are calculated based on the environment model (atmosphere and gravity models) according to the aircraft\u2019s position in the world"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000662_j.jmatprotec.2007.11.038-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000662_j.jmatprotec.2007.11.038-Figure4-1.png",
+        "caption": "Fig. 4 \u2013 Tooth flanks of a toothed wheel obtained with",
+        "texts": [
+            " Shift c corresponds to the value of the axial feed per wheel turn. A computer simulation of gear generation by chiselling was carried out in a similar way (Fig. 3b). The tool in this case is a solid model of a pinion cutter. The turn of the e c h n o l o g y 2 0 4 ( 2 0 0 8 ) 331\u2013342 tool by angle \u03d51 is related to the wheel turn \u03d52, in accordance with the transmission ratio of the technological gear. Examples of the flanks of the teeth obtained through hobbing and chiselling simulation is presented in Fig. 4. In the first case the flanks of the toothed wheel are composed of a number or regularly distributed surfaces corresponding to the successive layers of the material being removed by the tool blades. The surfaces are concave; therefore the tooth flanks are scaly in character. Their depth along the height of the teeth varies. The flanks of the teeth, chiselled by Fellows method, are also composed of a number of concave surfaces. They have concavities only along the profile along the tooth height (Fig. 4b). They run rectilinearly, tangentially to the helix. Taking the simulation parameters corresponding to the real generation parameters, it is possible to obtain a solid gear wheel model without the errors resulting from the real generation process (Dudley, 2002; Jehng, 2002; Michalski and Skoczylas, 2003). To obtain the models of the surfaces of the analysed gears teeth (Table 1), the following method of calculating the parameters of a computer simulation of gear generation was adopted: (1) For a hobbed wheel, the tool a hobbing cutter has 14 flutes on the perimeter"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003636_s00542-012-1544-7-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003636_s00542-012-1544-7-Figure4-1.png",
+        "caption": "Fig. 4 Comparison of pressure distributions between the Reynolds model and the Navier\u2013Stokes periodic model",
+        "texts": [
+            " This simulation used the thrust grooved bearing area with small gap of the Navier\u2013Stokes periodic model as shown in Fig. 3 which satisfies the Reynolds assumption. The pressure boundary conditions at the inlet and outlet are typically used to solve the Reynolds model. The boundary condition of atmospheric pressure was applied to inlet and outlet of the Reynolds model and the Navier\u2013Stokes periodic model only for the purpose of numerical verification. The simulated pressure distribution and load capacity of the Navier\u2013Stokes periodic model match well with those of the Reynolds model as shown in Fig. 4. The axial load capacity of the Navier\u2013Stokes periodic model (7.8 mN) matches well with that of the Reynolds model (7.6 mN). We also compared the characteristics of oil leakage and leak free operating speed of the Navier\u2013Stokes periodic model with the Bootsma and Tielemans\u2019 experimental results (Bootsma and Tielemans 1977). This example simulation used the Navier\u2013Stokes periodic model with the same seal design as the Bootsma and Tielemans\u2019 experimental model. Figure 5a, b shows the Bootsma and Tielemans\u2019 experiment and the simulation model of this research"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000373_50006-8-Figure6.23-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000373_50006-8-Figure6.23-1.png",
+        "caption": "Fig. 6.23. Single Point Turning Tool Geometries 4~",
+        "texts": [
+            "008 ipr Even further productivity gains, through higher cutting speeds, can often be achieved with cubic boron nitride (CBN) and ceramic tools, but the setup needs to be extremely rigid to preclude tool chipping. 42 CBN tools are often used when turning the harder nickel base (wrought and cast) and cobalt base cast alloys. Ceramic cutting tools allow speeds of 500-700 sfm with feeds of about 80% those of carbide. Typical ceramic tools include SiAION and SiC whisker reinforced A1203. The recommended geometries for single point turning tools are shown in Fig. 6.23. As mentioned earlier, it is important to use positive rake angles to insure that the metal is cut, rather than pushed. Positive rake angles also help to guide the chip away from the workpiece surface. The side and end relief angles, which provide clearance between the tool flank and the workpiece, are usually a compromise: too small of an angle will cause rubbing, while too large of an angle will provide insufficient cutting edge strength. The side cutting angle affects the load on the cutting edge and provides thickness and directional control to the chips"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003551_20130904-3-fr-2041.00119-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003551_20130904-3-fr-2041.00119-Figure4-1.png",
+        "caption": "Fig. 4. Topological interpretation of (48).",
+        "texts": [
+            " The property \u03b1\u2296 1 \u25e6 c\u03021\u03c31 \u25e6 \u03b1 \u2296 2 \u25e6 c\u03022\u03c32(s) < s, \u2200s\u2208R + \\ {0} (47) holds if and only if M(s) 6\u2265 0, \u2200s \u2208 R 2 + \\ {0} (48) is satisfied with ci = c\u0302i, i = 1, 2. The existence of c\u0302i > 1, i = 1, 2, satisfying (47) implies and is implied by the existence of c\u0304i > 1, i = 1, 2, satisfying (37). Indeed, property (47) guarantees (37) with c\u0304i = c\u0302i. The converse holds true with c\u0302i = c\u0304i/2. Hence, the obtuse angle problem (34) illustrated by Fig.3 is recast as (48). The condition (48) is given a topological interpretation in Fig.4. Property (48) ensures that the open set \u2126 = {s = [s1, s2] T \u2208 R 2 + :M(s) \u226a 0} (49) divides R 2 + \\ {0} into two disjoint sets. Under the map M : R 2 + \u2192 R 2, the set \u2126 is the preimage of the open negative orthant in R 2, i.e., the third quadrant in Fig.3. The boundary of \u2126 is given by the two curves li: \u03b1i(si) = ci\u03c3i,3\u2212i(s3\u2212i), i = 1, 2. The curve li is identical with the s3\u2212i-axis if \u03c3i,3\u2212i = 0. The unboundedness of \u2126 in the s3\u2212k direction is equivalent to the ISS property of \u03a3k since the unboundedness is equivalent to \u03b1k(\u221e) \u2265 ck\u03c3k,k\u22123(\u221e). Figure 4 (b) illustrates the case where only \u03a32 is ISS and \u2126 is unbounded only in the s1 direction. The boundedness of \u2126 in the s2 direction allows \u03a31 to be non-ISS. It is interesting to notice that based on Fig.4 (a), Jiang et al. [1996] constructed the max-type Lyapunov function (17). We can choose the strictly increasing function \u03c1\u0304 so that the curve s1 = \u03c1\u0304(s2) is a subset of \u2126 connecting the origin and (\u221e,\u221e) in Fig. 4 (a), i.e., M([s1, \u03c1\u0304 \u22121(s1)] T ) \u226a 0, \u2200s1 \u2208 R+ \\ {0} . (50) The existence of such a curve is guaranteed if (48) is satisfied, provided that the two subsystems are ISS (Jiang et al. [1996], Dashkovskiy et al. [2010]). Here, to ensure that V in (17) is a radially unbounded function defined on R N , the function \u03c1\u0304 needs to be defined on R+ and unbounded. This implies that the set \u2126 must be unbounded in both s1 and s1 directions. Thus, the max-type Lyapunov function (17) requires both subsystems to be ISS (Ito et al",
+            " Next, consider a linear function \u03c1\u0304 and let \u03c1\u0306 > 0 denote its coefficient. Then (50) holds if and only if M\u0306 [1, 1/\u03c1\u0306\u22121]T < 0 holds for the constant \u03c1\u0306 > 0. The Perron-Frobenius theorem again yields that there exits such a constant \u03c1\u0306 > 0 if and only if the largest eigenvalue of M\u0306 is negative. A choice of [1, 1/\u03c1\u0306\u22121]T is a corresponding right eigenvector of M\u0306 . Thus, if M is linear, a solution to the obtuse angle problem (34) in Fig.3 is a left eigenvector, while a solution to the separation problem (50) in Fig.4 (a) is a right eigenvector associated with M (Dashkovskiy et al. [2011]). The linearity of M implies ISS of the two subsystems. Notice that the left eigenvector remains a solution to (34) even when M(s) = M\u0306 [m1(s1),m2(s2)] T holds for some m1, m2 \u2208 P. Thus, the left eigenvector approach (34) is effective even if subsystems are not ISS. The interconnection (25) becomes a cascade system when \u03c32 = 0. According to Theorem 13, under Assumption 9, an interconnection of two iISS subsystems is always GAS"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002683_j.finel.2010.08.001-Figure20-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002683_j.finel.2010.08.001-Figure20-1.png",
+        "caption": "Fig. 20. Cylindrical shell: geometry and loading definitions.",
+        "texts": [
+            " Deformed configurations of the structure at the marked equilibrium points are depicted in Fig. 18. For the dynamical case, again in Fig. 17 the modejump at the first limit point can be observed while in Fig. 19 the deformed configurations at the marked instants are reported. A cylindrical shell of constant thickness and deformed by an applied compressive load is analysed. We consider vanishing radial and tangential displacements on both ends and E\u00bc3103, n\u00bc 0:3, as material parameters. Geometric parameters and problem definitions are given in Fig. 20. A 8 8 mesh for the symmetric quarter of the shell was considered. Then, v\u00bc0 for the nodes along the symmetric circumferential edge and u\u00bc0 for the nodes along the symmetric longitudinal edge. As before, proper treatment of the central points of the elemental edge at boundaries is carried out. Two cases, which differ for the values of radius R, thickness h and mass density r, were analysed. The computed equilibrium curves are displayed by external load parameter l and vertical deflection at the central point of the shell wc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002948_s00170-011-3420-5-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002948_s00170-011-3420-5-Figure1-1.png",
+        "caption": "Fig. 1 Aluminum die casting (material brand: ZL102)",
+        "texts": [
+            " Section 3 shows the experiment parameters, method of biomimetic laser-remelting process, and the performance of the diecasting die under actual production conditions. Section 4 illustrates the microstructure of the unit. Section 5 describes the application of biomimetic laser-remelting process on the succeeded die-casting die. Section 6 gives conclusions. Due to the high cost of die-casting die, one die casting and the corresponding die were selected elaborately. 2.1 The characteristics of the selected aluminum die casting The selected aluminum casting was produced by highpressure die casting, as shown in Fig. 1, called cover, which is used in vehicles. The material of the casting is ZL102. Though the geometry of the casting is not very complex, the dimension accuracy and the surface roughness are required strictly. The inner surface of die casting is required to keep the original die-casting surface. There are two platforms which have flatness checking requirements. The outer surface of the casting is cleaned by shot blast and then sprayed with black paint, as depicted in Fig. 1. The average wall thickness of the die casting is about 7 mm, which is thicker than general castings. Moreover, the wall thickness is not even. In the two-platform region, the max thickness reaches 18 mm, which create areas of high temperatures during solidification, the so called hot spots. Furthermore, there are sharp angles or edges near the ribs on outer surface and platforms on inner surface, which are known to promote or increase the risk of soldering [9] and corner cracking [6]. The strict requirements, uneven wall thickness, and corners in small radius lead to great difficulties in die-casting production and short service life of the die-casting die"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001064_s11012-009-9220-4-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001064_s11012-009-9220-4-Figure9-1.png",
+        "caption": "Fig. 9 Work piece before cutting of gear teeth",
+        "texts": [
+            " The emergence of new heat-resistant and high strength plastic materials has initiated improvements in gear manufacturing methods. This includes production of gears with the die method. In this method, molten or powdered material is poured into a specially designed mould having the shape of the final required gear. Examples of die forming processes are injection moulding, the powdered metal process, and forging [19]. With the above mentioned manufacturing methods, it is possible to manufacture gears whose teeth widths have been modified. In cutting methods, a gear is machined on a turning machine, as seen in Fig. 9, in a way resembling the load distribution on the gear surface, and then gear teeth are cut with gear teeth cutting methods. In the injection method, the gear is obtained with the help of dies, as explained before. 3.4 Hertzian stress along contact path Gear teeth are subjected to Hertz contact stresses and various surface damages. Overloading is one of the basic factors that cause abrasion, and scoring. Generally, a good correlation has been observed between spur gear surface fatigue failure and the computed elastic surface stress (Hertz stress) [20]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000456_j.triboint.2007.09.007-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000456_j.triboint.2007.09.007-Figure2-1.png",
+        "caption": "Fig. 2. Cross-section of couple-stress fluid film journal bearing.",
+        "texts": [
+            " In this figure, Om is the center of gravity of the rotor; O1, O2 and O3 are the geometric centers of the bearing, the rotor and the journal, respectively; 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 are the stiffness coefficients of the springs supporting the two bearing housings; C1 is the damping coefficient of the supported structure; C2 is the viscous damping of the rotor disk; r is the mass eccentricity of the rotor; f is the rotational angle; R is the inner radius of the bearing housing and r is the radius of the shaft. Fig. 2 illustrates the cross-section of the fluid film journal bearing. Note that (X, Y) are fixed coordinates, while (e, j) are rotational coordinates, in which e is the offset of the journal center and j is the attitude angle of the rotor relative to the X-coordinate direction. Applying the principles of force equilibrium, the forces acting at the journal center O3 (X3, Y3) are given by Fy \u00bc f e sin j f j cos j \u00bc ks\u00f0Y 2 Y 3\u00de=2, (2) in which fe and fj are the viscous damping forces in the radial and tangential directions, respectively, and are given by f e \u00bc f r and f j \u00bc f t",
+            " Substituting Eqs. (13) and (14) into Eqs. (1) and (2), respectively, yields the corresponding values of Fx and Fy. Let a1 \u00bc mL3R=2c2, then b1 \u00bc Z p 0 sin y cos y \u00f01\u00fe cos y\u00de3 12\u00f0l \u00de2\u00f01\u00fe cos y\u00de \u00fe 24\u00f0l \u00de3 tanh \u00f01\u00fe cos y\u00de=2l dy, g1 \u00bc Z p 0 cos2 y \u00f01\u00fe cos y\u00de3 12\u00f0l \u00de2\u00f01\u00fe cos y\u00de \u00fe 24\u00f0l \u00de3 tanh \u00f01\u00fe cos y\u00de=2l dy, d1 \u00bc Z p 0 sin2 y \u00f01\u00fe cos y\u00de3 12\u00f0l \u00de2\u00f01\u00fe cos y\u00de \u00fe 24\u00f0l \u00de3 tanh \u00f01\u00fe cos y\u00de=2l dy. Note the use of p as the upper integration limit in the expressions for b1, g1, and d1. From Eqs. (1) and (2) and Fig. 2, it can be shown that Fx and Fy may be rewritten as Fx \u00bc a1\u00bd\u00f0o 2 _j\u00de b1 2_ g1 cos j a1\u00bd\u00f0o 2 _j\u00de d1 2_ b1 sin j \u00bc cKp\u00f0x2 x1 cos j\u00de 2 , (15) Fy \u00bc a1\u00bd\u00f0o 2 _j\u00de b1 2_ g1 sin j\u00fe a1\u00bd\u00f0o 2 _j\u00de d1 2_ b1 cos j \u00bc cKp\u00f0y2 y1 sin j\u00de 2 . (16) From Eqs. (15) and (16), _ and _j are given by b1cKp\u00bd\u00f0y2 y1 sin j\u00de cos j \u00f0x2 x1 cos j\u00de sin j _ \u00bc d1cKp\u00bd\u00f0x2 x1 cos j\u00de cos j\u00fe \u00f0y2 y1 sin j\u00de sin j 4a1\u00f0g1d1 b21\u00de , \u00f017\u00de _j \u00bc o 2 cKp\u00bd\u00f0y2 y1 sin j\u00de cos j \u00f0x2 x1 cos j\u00de sin j 4a1d1 b21cKp\u00bd\u00f0y2 y1 sin j\u00de cos j \u00f0x2 x1 cos j\u00de sin j b1d1cKp\u00bd\u00f0x2 x1 cos j\u00de cos j\u00fe \u00f0y2 y1 sin j\u00de sin j 4a1 d1\u00f0g1d1 b21\u00de "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000784_acc.2008.4587070-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000784_acc.2008.4587070-Figure3-1.png",
+        "caption": "Fig. 3. Formation Parameters of Three Neighboring Helicopters",
+        "texts": [
+            " Using the elements of the vector p12, three geometrical parameters of the l \u2212 \u03b1 formation scheme are defined as follow: l12 = \u221a p2 12x + p2 12y \u03b112 = arctan(p12y /p12x ) (9) z12 = p12z \u03c82 = \u03c82 (10) In Eq. (10), we have also defined the yaw angle of the follower helicopter as our fourth control parameter. This results in a balanced input-output relation, where the dimension of the control input vector is equal to the number of control output parameters. As a result, the l \u2212 \u03b1 formation scheme output is assembled in the following form: y = g(x) = [ l12 \u03b112 z12 \u03c82 ]T (11) Fig. 3 shows the configuration of the l \u2212 l scheme. As for the l \u2212 \u03b1 case, here, we constrain the control point of one helicopter with respect to the control points of two other neighbors. In order to keep the ability of each agents to link with two different schemes simultaneously, the definition of the control point has been kept consistent for both cases. The following relation is held for point p3, which is the control point of the follower: pc1 + l13 + d1 = pc3 + d3 (12) Here, the vector l13 is the relative distance between the control points of the follower and the leader 3"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001632_2010-01-0228-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001632_2010-01-0228-Figure2-1.png",
+        "caption": "Fig. 2. Substructure A: free-free BIW",
+        "texts": [
+            " This approach provides a tool where the global frequency characteristics of a coupled structure are described in terms of component frequency response function matrices. (1) Equation (1) indicates how the system matrix of the coupled system HC can be calculated from the system matrices of the components (HA and HB) The indices are related to input, output and coupling points as indicated in Fig. 1. The FBS technique can be used to convert the free-free system to a constrained system: \u2022 subsystem A: free-free FRF's of BIW as shown in Fig. 2 \u2022 subsystem B: ground To represent the static test bench condition for torsional stiffness, the rear domes will be grounded (Fig. 3). To represent the static test bench condition for bending stiffness, rear and front domes will be grounded (Fig. 4). Forced Response The bending stiffness of a vehicle body is measured by clamping the body at the four domes and applying a load at the 4 seat bolting positions (F1 to F4) as is represented in Fig. 5. (2) After calculating the coupled system matrix Hc from equation (1) the displacement in the output points d(\u03c9) can be easily calculated using a forced response described in equation (2)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.26-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.26-1.png",
+        "caption": "Fig. 3.26. Velocity and force manipulability ellipses for a 3-link planar arm in a typical configuration for a task of controlling force and velocity",
+        "texts": [
+            " Restricting the analysis to a two-dimensional task space (the direction vertical to the surface and the direction of the line of writing), one has to achieve fine control of the vertical force (the pressure of the pen on the paper) and of the horizontal velocity (to write in good calligraphy). As a consequence, the force manipulability ellipse tends to be oriented horizontally for correct task execution. Correspondingly, the velocity manipulability ellipse tends to be oriented vertically in perfect agreement with the task requirement. In this case, from Fig. 3.26 the typical configuration of the human arm when writing can be recognized. An opposite example to the previous one is that of the human arm when throwing a weight in the horizontal direction. In fact, now it is necessary to actuate a large vertical force (to sustain the weight) and a large horizontal velocity (to throw the load for a considerable distance). Unlike the above, the force (velocity) manipulability ellipse tends to be oriented vertically (horizontally) to successfully execute the task"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003548_gt2013-94834-Figure11-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003548_gt2013-94834-Figure11-1.png",
+        "caption": "FIGURE 11: 2D (LEFT) AND 3D (RIGHT) MODELLED GEOMETRY",
+        "texts": [
+            " The effects of the rate of deformation could also be taken into account, for example by considering the transient temperature development. In case of periodic (eccentric) rubbing, the temperature increase and hence the compressive stress will oscillate, which may lead to a decrease in material strength due to fatigue. In order to validate the analytical results for the occurring thermo-mechanical stresses during labyrinth seal fin rubbing, various numerical FE studies were performed in Abaqus. Both 2D axisymmetric and 3D simulations with coupled temperaturedisplacement elements were run as depicted in Figure 11. For the 3D model only a segment with periodic boundary conditions was modeled in order to reduce computation time. Initial FE studies had shown that it is sufficient to model only a section of the labyrinth seal fin carrier with one seal fin. This is true as long as the modeled section is large enough compared to the seal fin size. Various seal fin geometries with and without fin inclination and tapering were studied. The initial temperature of the seal was set homogeneously to T0, which was varied between 500K and 900K",
+            " The yield surface was defined by the von-Mises stress and hardening effects were taken into account. The rub interaction with the stator was modeled by assuming a time-constant, axisymmetric surface friction heat flux and contact pressure into the seal fin contact surface. Various configurations with a concentrated heat flux were studied in order to simulate a \u201chot spot\u201d with varying width and position, as depicted in Figure 12. In a baseline configuration the heat flux was assumed constant over the whole width of the seal fin as indicated in Figure 11. In each of these configurations the total sum of incurred heat flux was kept identical by multiplying the baseline surface heat flux with the ratio of fin width over hot spot width. The various configurations allowed simulation of various types of temperature profiles with large axial and radial temperature gradients. In addition, different values for the nominal surface heat flux were modeled as part of a sensitivity study. Overall, about 40 FE simulations with different parameter combinations were run"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003977_j.proeng.2012.09.561-Figure14-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003977_j.proeng.2012.09.561-Figure14-1.png",
+        "caption": "Fig. 14 The fifth modal shape in Solid Works",
+        "texts": [
+            " The place of a fixation was realized by holders fixed to the stand (fig.4b). The material of the pipeline was predominantly zinc 99,995%, alloyed by copper and titanium. The aim of the measurement was to find out natural frequencies and modal shapes of pipeline model for two cases. The first case when springs of elastic support were not preloaded and when they were. Measurement results comparison acquired by Pulse6 for the case when springs were not preload with the results acquired by software Solid Works are presented on next figures (fig.5 - fig.14). Preload of pipeline support springs was created in one half of their length and the second half conserved its previous stiffness. Modal shapes did not differ for both investigated cases but differ in natural frequencies. From this reason we present only natural shapes of the first case, without preload. As we can see from the images there is from fourth modal shape also a deformation of pipeline cross-section and not only distortion of its shape. Corresponding natural frequencies of the particular modal shapes are presented in a table 1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001402_j.jmmm.2008.04.081-Figure10-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001402_j.jmmm.2008.04.081-Figure10-1.png",
+        "caption": "Fig. 10. The industrial 160 kW water-cooled RSW: transformer primary terminals (1), transformer (2), diodes housing (3), Rogowsky coil (4), water-cooling (5) and testing burden (6).",
+        "texts": [
+            " It can be avoided by the correct positioning of full-wave output rectifier diodes using specially chosen characteristics (paired diodes). Thus, any disturbing current spikes in the primary coil current of a highly loaded transformer can be avoided [6], as can be seen in Fig. 8 (calculation) and Fig. 9 (measurement). Currents are measured by the symmetrical RSW system \u2014 steady-state results. Rogowsky coil [11], while the transformer flux density B is measured by an additional measuring coil [5]. The water-cooled transformer with the output rectifier diodes of the industrial 160 kW RSW system used in testing is shown in Fig. 10. The analysis of RSW system has shown, that the current spikes in the transformer\u2019s primary current during the steady state operation occurs as a consequence of magnetic saturation in the iron core. Two main reasons for magnetic saturation of iron core and current spikes are the different ohmic resistances of transformer\u2019s secondary windings and different characteristics of the output rectifier diodes. Current spikes can be easily and efficiently eliminated by the correct positioning of carefully chosen diodes in the full-wave output rectifier"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002774_s12206-011-1203-4-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002774_s12206-011-1203-4-Figure3-1.png",
+        "caption": "Fig. 3. Stress cloud charts of the bolts with different loads.",
+        "texts": [
+            " Therefore, the stress of the bolt in each combination of factors is computed under three kinds of loads, respectively. The loads include tension, compression, and preload. Comparing the stress values of the bolt under the three kinds of the loads, the maximum stress value with each combination of factors can be identified, and this value is an indicator for the bolt connection performance. Alternate stress is also an important indicator. The bolt stress cloud charts under the three kinds of loads are shown in Fig. 3. From these figures, we can see the change in stress distribution when the bolt is supporting the tension and compression. It follows that the bolt is exposed to different alternate stress when slew bear- ings are rotating. Therefore, besides the computation of the bolt maximum stress, the alternate stress applied to the bolt when slewing bearings are working should be considered as an indicator. The maximum stress value is the primary indicator to be considered, and the alternate stress is the secondary indicator"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001361_icicic.2008.648-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001361_icicic.2008.648-Figure2-1.png",
+        "caption": "Figure 2. The free-body diagram of a RUAV.",
+        "texts": [
+            " The SIA-Heli-90 RUAV system is equipped with sensors including Inertial Measure Unit (IMU), GPS, digital compass, to obtain above accurate information about the motion of the helicopter in association with environmental information. It supports the wireless network communication with ground flight control station. The primary parameters are shown in table 1. RUAV dynamics obey the Newton-Euler equation for rigid body in translational and rotational motion. Here we consider a typical rigid RUAV in/near hover flight and the dynamic equation is conveniently described with respect to the body coordinate system, which is written as: The free body diagram of helicopter with respect to body coordinate system is as shown in figure 2. By employing the lumped-parameter approach, which considers the RUAV as the composition of the main rotor, tail rotor, fuselage, horizontal stabilizer, and vertical stabilizer. These components are considered as the source of forces and moments. The external force and moment in hovering can be written as: FB ext = \u23a1 \u23a3 XM YM + YT ZM \u23a4 \u23a6 + RTP\u2192B \u23a1 \u23a3 0 0 mg \u23a4 \u23a6 (2) MB ext = \u23a1 \u23a3 RM + YMhM + ZMYM + YT hT MM + MT \u2212XMhM + ZM lM NM \u2212 YM lM \u2212 YT lT \u23a4 \u23a6 (3) The forces and torques generated by the main rotor are controlled by TM , a1 and b1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000373_50006-8-Figure6.22-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000373_50006-8-Figure6.22-1.png",
+        "caption": "Fig. 6.22. Typical Tool Wear Features 4~",
+        "texts": [
+            " For example, the iron-nickel based alloys, which resemble stainless steels, are easier to machine than the nickel and cobalt based alloys. In general, as the amount of alloying elements increases for higher temperature service, the alloy becomes more difficult to machine. These high temperature characteristics place cutting tools under tremendous heat, pressure, and abrasion, leading to rapid flank wear, crater wear, and tool notching at the tool nose and/or depth of cut region, as illustrated in Fig. 6.22. 249 Due to their high temperature strengths, superalloys remain hard and stiff at the cutting temperature, resulting in high cutting forces that promote chipping or deformation of the tool cutting edge. In addition, since superalloys retain a large percentage of their strength at elevated temperatures, more heat is generated in the shear zone resulting in greater tool wear than with most metals. Since the forces required to cut superalloys are about twice those required for alloy steels, tool geometry, tool strength, and rigidity are all important variables"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001115_j.triboint.2009.11.005-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001115_j.triboint.2009.11.005-Figure1-1.png",
+        "caption": "Fig. 1. Contact area at (a) macroscopic level and (b) microscopic level.",
+        "texts": [
+            " The most significant are Amonton\u2019s and Coulomb\u2019s friction laws [1\u20135], which basically state that friction force is proportional to vertical load (FN) and independent of the characteristics of the contact, yielding the expression of the classical Coulomb friction model which defines the friction coefficient m as the ratio of friction to normal forces (1): m\u00bc Ff FN \u00f01\u00de In contrast with other rigid materials, friction of rubbers is characterized by several macroscopical dependencies: contact pressure, relative sliding speed and temperature, which can be easily observed in tribometer tests as well as in results shown in literature [6\u20138]. Therefore, a general friction function for rubbers should be expressed in the form (2): m\u00bc m\u00f0P; v; T\u00de \u00f02\u00de The particular mechanical characteristics of rubbers influence their frictional behavior, as has been demonstrated by many authors [9\u201311]. Rubber has low elastic modulus and high elongation and, therefore, its microscopic contact area is large as the rubber adapts to the shape of the surface asperities of the countermaterial (Fig. 1), showing in consequence high friction in dry conditions [12,13]. In contrast, the main statement of ll rights reserved. Amonton\u2019s and Coulomb\u2019s model is that the friction coefficient is independent of the vertical load and the macroscopic area of contact, which in terms of micro-scale physics means that the ratio of real area of contact (Ar\u2014interaction between the asperities of the two bodies in contact) to apparent area of contact (Ap\u2014macroscopic area) remains constant when the vertical load increases"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002886_1548512911414951-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002886_1548512911414951-Figure1-1.png",
+        "caption": "Figure 1. Conceptual design of Tiltrotor UAV.",
+        "texts": [
+            " Unlike its manned predecessors, its small size and capability of transitions between flight modes through tilting the body itself characterizes the unique nature of Tiltrotor UAV; this allows faster transitions and higher levels of stability in the helicopter mode. In general, airfoil dynamics are utilized up to the stall angle in modeling airplanes and helicopters. Owing to Tiltrotor UAV\u2019s genuine conversion method, a lookup table method based on instantaneous Reynolds number is asserted in modeling airfoil out of the stall angles. The conceptual design of Tiltrotor UAV (Figure 1) is realized according to the available airplane and helicopter models, with the purpose of combining the desirable and eliminating the undesirable features. The design phase is accomplished through the following steps. - The fuselage is the main component of the Tiltrotor UAV, holding all parts together with payloads, avionics and electrical systems. - Wings are fixed symmetrically at the sides of the fuselage, providing lift at forward speeds. A dihedral design is used through the wingtips in order to increase stability of the rolling motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002999_012-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002999_012-Figure1-1.png",
+        "caption": "Figure 1. Position vectors relating to a moving point Q.",
+        "texts": [
+            " The four ways of attacking rolling as a rotation about the instantaneous point of contact are finally illustrated by applying them successively to rolling of a semicircular hoop as a case presented in a recent article [1], where Turner and Turner disprove the above-mentioned misunderstanding, their case being a counterexample. When considering rolling as a rotation about the instantaneous point of contact as the Berkeley Physics Course and Ohanian actually do and how I was taught in my introductory course in mechanics, the angular momentum inserted in the torque rule in general terms is LQI = \u2211 i (ri \u2212 rQ) \u00d7 mivi; vi = dri/dt, (1) where, as indicated in figure 1, ri are the position vectors in an inertial frame of reference of the mass points mi of the rolling body, the point of contact is called Q and rQ is the position of this point in the inertial frame of reference. The index Q refers to the point about which moments are taken. The index I tells us that the velocities in the definition of LQI are evaluated in an inertial frame of reference. In the textbooks LQI is normally expressed as IQ\u03c9. The logical mistake then appears when, before differentiating, IQ is calculated with Q understood as the point of contact at any time, which means that rQ therefore should be considered a function of time and not a fixed point as assumed when using the normal torque rule"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002565_gt2011-46492-Figure12-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002565_gt2011-46492-Figure12-1.png",
+        "caption": "Figure 12. Design of Experiments: Input Variables 3 axial rows (Z-direction) displayed",
+        "texts": [
+            " Net blow-down force and tip reaction force due to blow-down were examined in the parametric study. The methodology previously outlined was used as the basis of a parametric study, to determine the effect of brush seal geometry on indicative brush seal performance parameters. This study was intended to determine the sensitivity of the outputs to the geometric inputs, and therefore the emphasis was on their variation rather than their absolute quantities. The six geometric inputs for an idealised brush seal domain are given in Fig. 12, and a proven method of Design of Experiments [18] was used to analyze their effect on the pre-defined outputs. A statistical approach was employed such that the minimum number of CFD simulations was required to ensure full coverage of the design space. A fractional 2-level factorial design was created for the six input variables, and this is given in Table 1. The number of runs was thus reduced from 64 (26) to 16 combinations of the input variables. A check for non-linearity in the output response was performed by identifying centre points mid-way between the low and high factor levels, and running additional simulations for this geometry. The result for simulation 15 was obtained from the averaged outputs of separate runs with four and five rows of bristles. 6 Copyright \u00a9 2011 by Rolls-Royce plc Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/gt2011/70340/ on 03/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use The assumed bristle tip heat flux is indicated in Fig. 12. In reality the value of this heat flux would directly relate to tip reaction forces, and thus would partly be dependent on bristle aerodynamic loading. However in this study, this flux was fixed to indicate how bristle pack effectiveness as a heat exchanger varies with geometry, as reflected by the tip temperature. Tip reaction and blow-down forces were also examined to give insight into contact forces and the brush seal\u2019s potential for frictional heat generation. Two extended bristle length simulations (18 and 19) were run to ensure that the range for L was sensible, and that no physical effects were missed"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002214_indcon.2009.5409419-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002214_indcon.2009.5409419-Figure1-1.png",
+        "caption": "Fig. 1 Mobile Robot Configuration",
+        "texts": [
+            " Assume that both geometry of A, B1, B2\u2026Bq and the locations of Bi\u2019s in W are accurately known. Assuming that no Kinematics Constraints limit the motion of A.. Given an initial position and orientation; a goal position and orientation of A in W, generate a path T specifying a sequence of positions and orientations of A avoiding contact with Bi\u2019s, starting at the initial position and orientation; and terminating at the goal position and orientation. We consider two dimensional workspace for mobile robot as shown in Fig.1. Robot is having initial coordinates xcoordinate (xo) and y-coordinate (yo). Similarly, target position coordinates are denoted as xt and yt respectively. Mobile robot\u2019s current position (calculated and updated at each step) 978-1-4244-4859-3/09/$25.00 \u00a92009 Can be denoted as xcurr and ycurr, Angle between target with respect to positive y axis is \u03b8tr. Robots pose (head) with respect to positive y axis is considered as \u03b8hr, \u03b8head is the heading angle between target and robot current position, Span(S) is the distance between left and right wheel, Vl and Vr are mobile robots left wheel and right wheel velocities, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001104_09544062jmes817-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001104_09544062jmes817-Figure4-1.png",
+        "caption": "Fig. 4 Rigid rotor model supported by symmetrical angular ball bearings",
+        "texts": [
+            " Then, the nonlinear bearing forces can be expressed as {Q} = \u23a7\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23a9 Qx Qy Qz My Mz \u23ab\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23ad = \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 N\u2211 i=1 (Q2i sin \u03b12i + F2i cos \u03b12i) N\u2211 i=1 (Q2i cos \u03b12i \u2212 F2i sin \u03b12i) sin \u03c6i N\u2211 i=1 (Q2i cos \u03b12i \u2212 F2i sin \u03b12i) cos \u03c6i N\u2211 i=1 [rr2(Q2i sin \u03b12i + F2i cos \u03b12i) \u2212f2DwF2i cos \u03b12i] cos \u03c6i N\u2211 i=1 [rr2(Q2i sin \u03b12i + F2i cos \u03b12i) \u2212f2DwF2i cos \u03b12i] sin \u03c6i \u23ab\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ad (23) rr2 = 0.5Dm + ( f2 \u2212 0.5)Dw cos \u03b10 (24) JMES817 \u00a9 IMechE 2008 Proc. IMechE Vol. 222 Part C: J. Mechanical Engineering Science at MICHIGAN STATE UNIV LIBRARIES on June 15, 2015pic.sagepub.comDownloaded from A rigid rotor with a single disc model supported by symmetrical angular ball bearings is shown in Fig. 4; the equations of motion for the rotor bearing system may be written as follows M X\u03082 + CX\u03072 + 2Qx = FX M Y\u03082 + CY\u03072 + 2Qy = FY + me\u03c92 cos \u03c9t M Z\u03082 + CZ\u03072 + 2Qz = FZ \u2212 Mg + me\u03c92 sin \u03c9t Iy \u03b8\u0308y \u2212 Iz\u03c9\u03b8\u0307z + 2My = MY Iz \u03b8\u0308z + Iy\u03c9\u03b8\u0307y + 2Mz = MZ \u23ab\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23ad (25) where me\u03c92 indicates the unbalance load due to the rotor mass eccentricity e. Using the Newmark-\u03b2 method, the differential equations of motion can be solved, and the transient responses at every time increment are obtained. Because the displacement of a bearing is equal to the displacement of a rotor, the dynamic characteristics of the ball bearing are impacted by vibration of the rotor system"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure2-1.png",
+        "caption": "Fig. 2 Coordinate system",
+        "texts": [
+            "org/ on 01/28/201 ok min = arcsin rt sin ok \u2212 m rt , ok max = arcsin rt sin ok + m rt where the parameters ik and ok are the surface coordinates of the circular-arc blade edges, the subscripts l and r represents the left and right cutter heads, respectively, ik and ok are the pressure angles of the inside and outside cutter blades, respectively, and m is the module of the work gear. 3 Mathematical Model of the Face-Hobbing Cutter Head Equations 1 and 2 define the cutting edge on the front cutting plane of the inner and outer cutter blades, which rotate about the head cutter axis and sweep out the generating surface of the head cutter. As shown in Fig. 2, the coordinate system Sb is rigidly attached to the head cutter with rotation axis Yb, and the vector of rotation axis Yb is inward into the face of the paper. The coordinate systems Sa and Sl are rigidly connected to the front cutting plane T of the cutter blade and the cutting edge, respectively, while Sm and Sn are the auxiliary coordinate systems that describe the relative position of the cutter edge on the cutter head. The front cutting plane T is tangent to the rolling circle, Hk is the rotation angle of the surface coordinates of the head cutter, the parameter k is the initial angle of the cutter blade, and the parameter rc= ObOm is the nominal head cutter radius, as defined in Fig. 2. Other parameters of the cutter heads include the offset angle H and the setup distance of cutting edge EHj. The subscript of the cutter blade s for the head cutter Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use t b w a 4 t s w e t b w J Downloaded Fr j may be replaced with i or o for the inner or outer blade, respecively. The surface geometry of the head cutter can be represented y the following matrix transformation equation: Rb jk, Hk = Mbn Hk \u00b7 Mnm \u00b7 Mml \u00b7 Mla \u00b7 Ra sj jk 3 here Mla = 1 0 0 EHj 0 1 0 0 0 0 1 0 0 0 0 1 Mml = cos H 0 \u2212 sin H 0 0 1 0 0 sin H 0 cos H 0 0 0 0 1 0 0 0 1 ournal of Mechanical Design om: http://mechanicaldesign"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003031_iros.2011.6094957-Figure9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003031_iros.2011.6094957-Figure9-1.png",
+        "caption": "Fig. 9. Graspability maps for a coffee cup and: a) A Barrett hand; b) A DLR hand II; c) A DLR-HIT hand II.",
+        "texts": [
+            " The graspability map is also a useful tool to compare the grasp capabilities of different mechanical hands. For instance, Fig. 7 show the graspability maps for the banana using a 3-fingered Barrett hand, a 4-fingered DLR hand II, and a 5-fingered DLR-HIT hand II [28]. In this case, the potential poses of the set \u0393 are generated with 50 points defined on the sphere inside each voxel, and 12 possible orientations per point (i.e. 600 possible poses per voxel). Fig. 8 filters the previous results, to show only the spheres which provide the higher number of valid poses to get an FC grasp. Fig. 9 shows the graspability maps for the coffee cup using the same set \u0393 with 600 poses per voxel. When comparing the maps for all the hands, the main highlight is that the graspability maps for the 4-fingered hand are much larger than the other two maps. In fact, that hand has a larger size, as shown in Fig. 10, and has a larger workspace which allows more potential poses to get FC grasps on different objects. For instance, the graspability maps for the banana shown in Fig. 7 contain 11,006 valid poses for the Barrett hand, 44,425 for the DLR hand II and 3,687 for the DLR-HIT hand II. In the case of the cofee cup (Fig. 9), although the 4- and 5-fingered hand are both anthropomorphic, the graspability map for the 5-fingered hand contains most of its points above the cup and towards the handle side, i.e. this hand is able to grasp the cup using basically the handle or the internal surface of the cup. In this paper we have proposed an algorithm for the offline computation of the graspability map, a representation of the poses for a mechanical hand that might lead to a precision force closure grasp. The map is specific for an object and hand"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002424_med.2012.6265709-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002424_med.2012.6265709-Figure2-1.png",
+        "caption": "Fig. 2. Helicopter Configuration",
+        "texts": [],
+        "surrounding_texts": [
+            "I. INTRODUCTION\nN fuzzy control literature, fuzzy PID controllers (FPID) are presented as an alternative to overcome the lack of the classical PID controllers when used with nonlinear systems and/or complex systems, in particular interconnected systems [1-7]. However, the challenge to take up is how to design these controllers considering the big number of parameters to be tuned compared to the classical PID controllers which have only three parameters that can be tuned easily by using the available rules in literature [8].\nNowadays, the heuristic optimization has been gained a special interest to solve the FPID controllers design problem. Among these methods we can find, genetic algorithms (GA) [9-12], ant colony optimization (ACO) [13] and particle swarm optimization (PSO) [12], [14].\nThe above listed methods are used for single input and single output (SISO) systems. In this paper, we propose a PSO tuning method to design decentralized typical fuzzy PI (FPI) controllers applied for the stabilization of a helicopter model, which is essentially nonlinear, with two inputs and two outputs (TITO), and open loop instable. The performance of the proposed method is evaluated by a fitness function associated with the whole control system performance. Without loss of generality, we use a square error fitness function (SE) to quantify the quality of two step responses of closed-loop system.\nWe consider in this paper, the typical FPI controllers presented in [1]. This type of FPI can be linear or nonlinear depending on the membership functions positions. The linear FPI controller is obtained by modal positions of the\nManuscript received January 26, 2012. H. Boubertakh is with LAJ-Laboratoire d\u2019Automatique de Jijel, FST, University of Jijel, BP. 98, Ouled Aissa, 18000, Jijel, Algeria (email : boubert_hamid@yahoo.com)\nS. Labiod is with LAJ, FST, University of Jijel, BP. 98, Ouled Aissa, 18000, Jijel, Algeria (email: labiod_salim@yahoo.fr)\nM .Tadjine is with LCP, Ecole Nationale Polytechnique, 10 av. Hassen Badi, BP.182, El Harrach, Algiers, Algeria. (email : tadjine@yahoo.fr)\nmembership functions. However, a nonlinear behavior can be reached by handling these quantities from their modal positions or to use some existing optimization methods. Thus, these typical FPI controllers can deal with complex and nonlinear processes and combine the advantages of classical and fuzzy control namely; simplicity, nonlinearity and interpretability. Moreover, the tuning time will be reduced by the exploiting the available knowledge and the human operator experience with the plant under control.\nThe remainder of this paper is organized as follows. Section II presents the design settings of typical FPI controllers. Section III presents the helicopter model. Section IV gives the decentralized control structure. Section V gives a brief overview the PSO and presents the proposed tuning method. The simulation results are given in section VI. Section VII concludes de the paper.\nII. THE FPI CONTROLLERS\nThe incremental FPI control law is given by 1 \u2206 (1) \u2206 , (2) The error and the change of error are defined as\n(3) \u2206 1 (4)\nWhere f is the function of the fuzzy inference system, is the control signal, \u2206 is the change of error is the reference signal and is the controlled output. According to the typical structural design of FPI controllers in [1], we use FPI controllers with 3 membership functions for input variables and 5 singletons for the output variable as shown on figure 1. Now, unless otherwise specified, we consider that the input membership functions and output singletons are evenly distributed on symmetrical universes of discourse (i.e. 2 ). Moreover, and to avoid saturation we consider that the universes of discourse are large enough such that all possible inputs are always within their limits. The rule base of the controllers is donated by table 1. The crisp change of control action is determined by the average sum:\n\u2206\n(5)\nwhere is the conclusion of rule and is its truth value calculated by the algebraic product method given by\n. \u2206 (6)\nI\n978-1-4673-2531-8/12/$31.00 \u00a92012 IEEE 633",
+            "is the membership grade of the input variable\nevaluated in rule by the corresponding membership function.\nThe labels NB, NS, ZR, PS, and PB refer respectively to the linguistic terms; Negative Big, Negative Small, around Zero, Positive Small and Positive Big.\nThe inferred output is calculated by (5) and (6). After calculation [1], we find that:\n\u2206 (7)\nEquation (7) shows clearly that the typical FPI controllers are identical to the classical PI controllers with the following\nproportional and integral gains: and .\nWe note here that a nonlinear behavior of the above designed FPI controllers can be reached by moving the input membership functions and/or the singletons from their modal positions.\nWe can conclude that the FPI controllers presented here combine the advantages of classical PI controllers and fuzzy controllers; the simplicity, the physical interpretability, and the possible nonlinearity of their structures.\nIII. HELICOPTER MODEL\nThe CE150 Helicopter Model of Humusoft Ltd [15] consists of a body carrying two DC motors. These motors drive the propellers. The body has two degrees of freedom. The axes of the body rotation are perpendicular as well as the axes of the motors. Both body position angles, i.e. azimuth angle  in horizontal and elevation angle \u03a6 in vertical plane are influenced by the rotating propellers simultaneously. The DC motors for driving propellers are controlled proportionally to the output signal of the computer. The helicopter model is a MIMO dynamical system with two manipulated inputs and , and two measured outputs \u03a6, . All inputs and outputs are coupled. The user of the simulator communicates with the\nsystem via the data-processing interface, the entries ( and ) and outputs are scaled in the interval [- 1,+1], where '' 1 '' is called Machine Unit MU . The mathematical model of the helicopter is given by the following differential equations system:\n0.8764 3.4325 0.4211\n0.0035 46.35 0.8076 0.0259 2.9749\n21.4010 31.8841 14.2029\n21.7150 1.4010 6.6667 2.7778 2\n4 8 4 2 4 1.3333 0.0625\nwhere:  ,  , \u03a6 , , and to\nare state variables representing the two DC motors and the coupling effects.\nIV. CONTROL STRUCTURE\nThe helicopter simulator can be seen like an interconnection of two subsystems; the elevation subsystem characterized by the input and the output  and the azimuth subsystem characterized by the input and the output .\nThe problem is to stabilize the system around a setpoint  , \u03a6 . For that we use decentralized control method; the two subsystems are controlled by two FPI controllers with structure design presented in section III. The control structure is presented on figure 3.\nj\n978-1-4673-2531-8/12/$31.00 \u00a92012 IEEE 634",
+            "V. TUNING THE FPI CONTROLLERS USING PSO\nA. Overview of the particle swarm optimization\nPSO is nature-inspired heuristic optimization method which first proposed by Kennedy and Eberhart [16]. It belongs to the category of Swarm Intelligence methods. Its development was based on mimicking the movement of individuals within a swarm (i.e., fishes, birds, and insects) in an effort to find the optima in the problem space. It has been noticed that members of the swarm seem to share information among them. This communication fact leads to increase efficiency of the swarm. The PSO algorithm searches in parallel using a group of individuals similar to other population-based heuristic optimization techniques. PSO technique conducts search using a population of particles, corresponding to individuals. Each particle represents a candidate solution to the problem at hand. In a PSO system, particles change their positions by \u201cflying\u201d around in a multidimensional search space. Particle in a swarm adjust its position in search space using its present velocity, own previous experience, and that of neighboring particles. Therefore, a particle makes use of best position encountered by itself and that of its neighbors to steer toward an optimal solution. The performance of each particle is measured using a predefined fitness function, which quantifies the performance of the optimization problem.\nThe mathematical expressions for velocity and position updates are given by (9) and (10) respectively.\n(9) (10)\nWhere: 1 , 1 , 1 . , and are respectively the number of particles in the swarm, the dimension of particle, and the maximum iterations.\nand are respectively the position the velocity of\nparticle in the dimension at iteration , is a personal best of particle in the dimension , is a global best of all particles in the dimension , is inertia weight factor , and are acceleration constants, and are random numbers in interval 0, 1 .\nB. The FPI Controllers tuning problem using PSO\nIn the proposed PSO tuning method for decentralized FPI controllers, the tuning vector is composed by the parameters of the two controllers.\nThe components of the tuning parameters vector can be defined according to the available knowledge and the control objective. In the aim to get nonlinear controllers, we consider the tuning parameters\u2019 vector composed by the positions of the membership functions over the input spaces and the positions of singletons over the output spaces for both the two controllers.\nTo simplify, we consider a symmetrical fuzzy PI controller such as the positive and negative parameters absolute values are equal. Thus, the parameter\u2019s vector is\nP E D U U E D U U T (10)\nIn order to implement the proposed method, we need the minimum and the maximum bounds for each tuning parameter. The initial parameters values are randomly distributed between these bounds. A best choice of these bounds can drastically reduce the tuning time. This is possible, since the FPI controllers have interpretable parameters, and can incorporate human operator knowledge about the system under control.\nA squared error fitness function (SE) (11) is used to quantify the effectiveness of a given FPI controller; it is evaluated at the end of a step-response of the closed-loop system under control.\n\u2211 \u2211 (11)\nN is the total number of samples, is the number of transient samples system output.\nThe fitness function defined in (11) takes into account the error over the time interval , . By setting to approximate the peak time of the system at which the step response of the closed-loop system reaches the corresponding first peak.\nIn the FPI controllers tuning problem using PSO, a particle is one value of parameters vector. A particle swarm is group of parameters vector values which are randomly initialized in search space defined by limits of the components of the tuning parameters vector. So, a the particle position is the FPI parameter value vector, the best parameters values vector regarding the fitness function, is the global best parameters values vector of all particles.\nThe proposed method is used to tune simultaneously the two decentralized FPI controller such as to minimize square error sum cost function (11) at the end of step-responses of the two subsystems of the helicopter model. In the simulation, we take the following parameters: the sample time 0.1 , the peak-time for the two subsystems 10 , and one step episode time is 50s.\nThe premise parameters bounds are given by table 2 and the initial rule base for the two FPI controllers is given by table 3.\n978-1-4673-2531-8/12/$31.00 \u00a92012 IEEE 635"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002022_1077546310384002-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002022_1077546310384002-Figure1-1.png",
+        "caption": "Figure 1. Unsymmetrical flexible rotor bearing system.",
+        "texts": [
+            ") nonlinear dynamic model is formulated and the gyroscope moment is taken into account for the rotation of the unsymmetrical rotor. A proportional\u2013integral (PI) controller is designed and added to the nonlinear equations of motion to adjust the displacement of the rotor inside of the bearing and consequently change the oil film coefficients (damping and stiffness). Numerical simulations are performed in order to suppressing efficiency of the control system in reducing vibration. Consider an unsymmetrical flexible rotor mounted in two tilting pad journal bearings as shown in Figure 1. The two bearings A, B are both actively lubricated by the two servo-valves. In Figure 1 \u00f0x, y, s\u00de are fixed coordinates, O0 is the center of gravity of the disk and with the eccentricity e. mT, mA, mB are the masses of the disk, TPJBs A and B respectively. l and a are the lengths of AB and AO. The gyroscope moment H appears along with the rotation of the unsymmetrical rotor. x and y are instantaneous rotations of the disk around x and y coordinates, xo, yo xA, yA xB, yB are the instantaneous horizontal and vertical displacements of the disk center and bearing A and B respectively. The equations of motion can be formulated as follows according to the rotor dynamics theories mA \u20acxA \u00fe 1 l \u00f0 bk11 \u00fe k41\u00dexO \u00fe 1 l 2 \u00f0b2k11 2bk14 \u00fe k44\u00dexA \u00fe 1 l 2 \u00bdabk11 \u00fe \u00f0b a\u00dek14 k44 xB \u00fe 1 l 2 \u00f0 bk14 \u00fe k44\u00de Oy \u00bc FAx \u00f01\u00de mA \u20acyA \u00fe 1 l \u00f0 bk22 \u00fe k32\u00de yO \u00fe 1 l 2 \u00f0b2k22 2bk23 \u00fe k33\u00de yA \u00fe 1 l 2 \u00bdabk22 \u00fe \u00f0b a\u00dek23 k33 yB \u00fe 1 l \u00f0bk23 k33\u00de Ox \u00bc FAy \u00f02\u00de mT \u20acxO \u00fe k11xO \u00fe 1 l \u00f0 bk11 \u00fe k14\u00dexA \u00fe 1 l \u00f0 ak11 k14\u00dexB \u00fe k14 Oy \u00bc mT 2e cos t \u00f03\u00de mT \u20acyO \u00fe k22yO \u00fe 1 l \u00f0 bk22 \u00fe k23\u00de yA 1 l \u00f0ak22 \u00fe k23\u00de yB k23 Ox \u00bc mT 2e sin t mTg \u00f04\u00de Jd \u20ac x \u00feH _ y k32yO \u00fe 1 l \u00f0bk32 k33\u00de yA \u00fe 1 l \u00f0ak32 \u00fe k33\u00de yB \u00fe k33 Ox \u00bc 0 \u00f05\u00de at East Tennessee State University on June 5, 2015jvc"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001082_s11071-007-9205-6-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001082_s11071-007-9205-6-Figure2-1.png",
+        "caption": "Fig. 2 Displacement field",
+        "texts": [
+            " The first and second Euler angles, \u03b1 and \u03b2, are the so-called elastic bending rotation angles at S about the third axis of FB and about the second axis of the updated FB by \u03b1, respectively. The third Euler angle, \u03b3 , is the so-called elastic twisting angle at S about the first axis of FS [7]. In some contexts these three Euler angles are called Bryant angles [2]. To avoid lengthy expressions, the beams with circular and square cross-sections are considered only. It implies that FS is a principal frame for the beam crosssection and the two moments of cross-sectional area about the second and third axes are equal as shown in expressions (6). In Fig. 2, \u03c3 is a general point of the beam cross-section. Position vector of \u03c3 from S projected onto FS is shown by p. It is a constant vector in 3D Euler\u2013Bernoulli beam theory and in this paper to neglect in-plane and out-of-plane warpings\u222b A pT pd A = 2J, [JS] = \u222b A \u2212 p\u0303 p\u0303d A = J \u23a1\u23a2\u23a32 0 0 0 1 0 0 0 1 \u23a4\u23a5\u23a6 , p = \u23a1\u23a2\u23a30 y\u0302 z\u0302 \u23a4\u23a5\u23a6 , p\u0303 = \u23a1\u23a2\u23a3 0 \u2212z\u0302 y\u0302 z\u0302 0 0 \u2212y\u0302 0 0 \u23a4\u23a5\u23a6 . (6) 3.1 Beam structural constraints Figure 3 displays two holonomic constraints among the six simply created elastic coordinates. By the application of these constraints, the two excess coordinates are eliminated",
+            " (18) 5 Variation of gravitational potential energy Variation of gravitational potential energy is given by the following equation [7]: \u03b4U g = \u222b L 0 \u222b A {[0 0 1]\u03b4\u03b7}g\u03c1d A ds = g\u03c1[0 0 1] \u222b L 0 \u222b A { \u03b4b + R IB [ \u03b4d + R BS \u03b4\u0303S p ] + R IB \u03b4\u03c0\u0303 B [ d + R BS p ]} d A ds = g\u03c1[0 0 1] \u222b L 0 ( \u03b4b \u222b A d A + R IB \u03b4d \u222b A d A + R IB R BS \u03b4\u0303S \u222b A pd A + R IB \u03b4\u03c0\u0303 Bd \u222b A d A + R IB \u03b4\u03c0\u0303 B R BS \u222b A pd A ) ds. Springer Since Sn is the center of cross-sectional area, it is simplified as follows: \u03b4U g = g\u03c1 A[0 0 1] \u222b L 0 { \u03b4b + R IB \u03b4d \u2212 R IB d\u0303\u03b4\u03c0 B } ds (19) 6 Variation of elastic potential energy In Fig. 2, \u03c3 is the general point of the beam\u2019s medium. Vector p is assumed to be constant, implying the lack of in-plane and out-of-plane cross-section warping in the beams [7]. The displacement field is = \u23a1\u23a2\u23a3 x y z \u23a4\u23a5\u23a6 = \u03be \u2212 \u03be | before elastic deformation = \u23a7\u23aa\u23a8\u23aa\u23a9 \u23a1\u23a2\u23a3u + s v w \u23a4\u23a5\u23a6 + R BS p \u23ab\u23aa\u23ac\u23aa\u23ad \u2212 \u23a7\u23aa\u23a8\u23aa\u23a9 \u23a1\u23a2\u23a3 s 0 0 \u23a4\u23a5\u23a6 + \u23a1\u23a2\u23a31 0 0 0 1 0 0 0 1 \u23a4\u23a5\u23a6 p \u23ab\u23aa\u23ac\u23aa\u23ad = \u23a1\u23a2\u23a3 u v w \u23a4\u23a5\u23a6 + R BS \u23a1\u23a2\u23a30 y\u0302 z\u0302 \u23a4\u23a5\u23a6 \u2212 \u23a1\u23a2\u23a30 y\u0302 z\u0302 \u23a4\u23a5\u23a6 . Linear part of Green\u2013Lagrange geometric strain ten- sor is taken into consideration as follows: \u03b5\u0302i j = 1 2 ( \u2202 i \u2202x j + \u2202 j \u2202xi ) , \u2202 \u2202s \u23a1\u23a2\u23a3 x y z \u23a4\u23a5\u23a6 = \u23a1\u23a2\u23a3u\u2032 v\u2032 w\u2032 \u23a4\u23a5\u23a6 + R BS \u23a1\u23a2\u23a3\u2212y\u0302\u03baz + z\u0302\u03bay \u2212z\u0302\u03bax y\u0302\u03bax \u23a4\u23a5\u23a6 , \u2202 \u2202y \u23a1\u23a2\u23a3 x y z \u23a4\u23a5\u23a6 = \u2202 \u2202z \u23a1\u23a2\u23a3 x y z \u23a4\u23a5\u23a6 = \u23a1\u23a2\u23a30 0 0 \u23a4\u23a5\u23a6 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002150_robot.2010.5509342-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002150_robot.2010.5509342-Figure1-1.png",
+        "caption": "Fig. 1. Principle of the osteotomy system",
+        "texts": [
+            " Section II describes the principle of the approach and introduces the notations used in the paper. Section III presents the control algorithm developed for static bones. It is a classical position-based visual servoing algorithm with slight modifications. The main contribution of this paper is the learning algorithm proposed in section IV for bones undergoing breathing motions. Experimental results are presented in section V. As mentionned in the introduction, we use the laser osteotomy setup developed by the university of Karlsruhe [2]. The principle of this approach is depicted on Fig. 1. A rigid 978-1-4244-5040-4/10/$26.00 \u00a92010 IEEE 4573 body constituted by optical markers is attached to the bone to be cut. In the following, we will refer to this rigid body by the name target. Another rigid body constituted by optical markers is attached to the robot\u2019s end-effector. The positions of the markers are measured by an optical tracking system. In this approach, a multi camera optical tracking system from ART with six cameras is used. The multi camera approach helps to avoid occlusion of the optical markers"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003147_1.4007806-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003147_1.4007806-Figure4-1.png",
+        "caption": "Fig. 4 Multiple defects model",
+        "texts": [
+            " Since the focus of this paper is on the ADFH, the Fourier series is considered and the mth spectral component of the response is written in terms of its magnitude and phase, as seen in Eq. (3) ym\u00f0t\u00de \u00bc Am sin\u00f02pmfdt\u00fe w\u00de (3) The relationship between the spectral component and the time and frequency domain is shown in Fig. 3, where ym is the mth spectral component, Am and w are the magnitude and phase, respectively, and fd is the defect characteristic frequency. 2.2.2 Multiple Faults. The rolling element strikes the first defect located in h \u00bc 0 when t \u00bc 0, as shown in Fig. 4. At time t \u00bc t0, in Eq. (4), the rolling element strikes the second defect located in h \u00bc h0 t0 \u00bc h0=\u00f02pfc\u00de (4) where fc is the frequency of the cage rotation. Each defect on the deterministic race has a defined characteristic defect frequency, which is related to the geometry of the bearing and the speed of the shaft rotation. Another impact at t0 of the same race causes the time shift [17] in Eq. (3) with the same characteristic defect frequency as the first defect. If it is assumed that the magnitude of the second defect response is equal to that of the first defect, the mth spectral component of the second defect is as follows ym\u00f0t t0\u00de \u00bc Am sin\u00f02pmfd\u00f0t t0\u00de \u00fe w\u00de \u00bc Am sin\u00f02pmfdt\u00fe w 2pmfdt0\u00de (5) The summation of the spectral component of the two defects is given in Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000070_978-1-4020-4941-5_42-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000070_978-1-4020-4941-5_42-Figure6-1.png",
+        "caption": "Figure 6. Actuated mechanism of family-3.",
+        "texts": [
+            " decouple the motion about the axes k0 and i1. Indeed, since a force is not able to generate moments about the lines it crosses, it is clear that every UPS-leg whose connecting joint on the base is centered in a point Bk, which lies on k0, makes it possible to control rotations about i1 only, while every UPS-leg whose connecting joint on the platform is centered in a point Pk, which lays on i1, makes it possible to control rotations about k0 only. A decoupled actuated manipulator obtained from a US-PM of family-3 is represented in Fig. 6 (UPS-legs are drawn as telescopic legs). The actuated UPS-leg, P5B5, controls the rotation about the axis k0 only, while the actuated UPS-leg, P6B6, controls the rotation about the axis i1 only. Note that the manipulator obtained from the mechanisms of family-3 coincides with the fully parallel spherical wrist with the P actuator on the leg P4B4 locked. 5 Kinematic, Workspace and Singularity Analyses Due to the decoupled actuation of the rotations of the mechanism about the k0 and i1 axes, the direct kinematic, workspace and singularity analyses are very straightforward"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003709_012082-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003709_012082-Figure6-1.png",
+        "caption": "Figure 6. LPM 34 DOF model [1]. t: Torsional",
+        "texts": [],
+        "surrounding_texts": [
+            "We have previously applied a number of reduction techniques to this problem, and found that the Craig-Bampton (CB) method [8] was most appropriate because it uses a combination of spatial coordinates defining interactions between the internals and the casing (where the interaction is nonlinear and time-varying), and modal coordinates for the rest giving a minimum total number of DOFs for a given frequency range. In this method, the CB modes of the structure are derived assuming that the master degrees of freedom are held fixed. It enables defining the frequency range of interest by retaining only the modes up to a defined upper limit. Note that the low order modes are greatly affected by the restraints applied to the master DOFs, but boundary conditions have little effect on the frequency of the high order modes defining the frequency range. The decomposition of the model into both physical DOFs (master DOFs) and modal coordinates allows the flexibility of connecting the finite elements to other substructures, while maintaining a reasonably good result within a specified frequency range. Our earlier studies had shown that the simple LPM gave good results at high frequencies, and the interaction problems were primarily in the mid frequency range."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0002441_i2mtc.2012.6229339-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002441_i2mtc.2012.6229339-Figure3-1.png",
+        "caption": "Figure 3. 3D model of the measuring bench.",
+        "texts": [
+            " More detailed analyses of a valve can be found in [6], [7] and in [10]-[12], [10] and [11] formulating FMEA (finite element analysis) and analytic models. In this paper, a specific pull type solenoid was studied and all presented results are related to it. Plunger position is taken such that it increases with the plunger being more immerged into the housing, so at the fully deenergized state it is zero. In order to analyze and measure a given solenoid actuator a dedicated measuring and testing environment has been built, illustrated in Fig. 2. Into a specific mechanical device, presented in Fig. 3, the tested solenoid can be mounted. The layout is arranged in a vertical configuration and the external load is provided by the gravitational force of certain masses. In steady state, this force is constant regardless of the plunger position. Regarding position measurement, it is achieved by transmissive optical sensors. These devices host a light source (diode) and photo detector (transistor) facing each other, operated at the infrared region. Since the amount of light arriving to the photo transistor is determined by the size of the aperture, position information is extracted from the emitter current. As it is shown in Fig. 3, the aperture is changed by an opaque plate moving with the plunger. For producing the necessary PWM waveform and acquiring the electrical and mechanical signals (e.g. current, position) of the solenoid, a dedicated HW (hardware) has been built hosting a 16 bit DSP (digital signal processor) (dsPIC33FJ128GP from Microchip). From a LabVIEW based PC (personal computer) interface the DSP can be accessed, from which the sampled data is transferred to LabVIEW for further high level signal processing"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000866_978-1-84628-642-1_3-Figure3.9-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000866_978-1-84628-642-1_3-Figure3.9-1.png",
+        "caption": "Fig. 3.9. Composition of elementary rotational velocities for computing angular velocity",
+        "texts": [
+            "62) where the analytical Jacobian JA(q) = \u2202k(q) \u2202q (3.63) is different from the geometric Jacobian J , since the end-effector angular velocity \u03c9e with respect to the base frame is not given by \u03c6\u0307e. It is possible to find the relationship between the angular velocity \u03c9e and the rotational velocity \u03c6\u0307e for a given set of orientation angles. For instance, consider the Euler angles ZYZ defined in Sect. 2.4.1; in Fig. 3.8, the vectors corresponding to the rotational velocities \u03d5\u0307, \u03d1\u0307, \u03c8\u0307 have been represented with reference to the current frame. Figure 3.9 illustrates how to compute the contributions of each rotational velocity to the components of angular velocity about the axes of the reference frame: \u2022 as a result of \u03d5\u0307: [\u03c9x \u03c9y \u03c9z ]T = \u03d5\u0307 [ 0 0 1 ]T \u2022 as a result of \u03d1\u0307: [\u03c9x \u03c9y \u03c9z ]T = \u03d1\u0307 [\u2212s\u03d5 c\u03d5 0 ]T \u2022 as a result of \u03c8\u0307: [\u03c9x \u03c9y \u03c9z ]T = \u03c8\u0307 [ c\u03d5s\u03d1 s\u03d5s\u03d1 c\u03d1 ]T , and then the equation relating the angular velocity \u03c9e to the time derivative of the Euler angles \u03c6\u0307e is9 \u03c9e = T (\u03c6e)\u03c6\u0307e, (3.64) where, in this case, T = \u23a1 \u23a3 0 \u2212s\u03d5 c\u03d5s\u03d1 0 c\u03d5 s\u03d5s\u03d1 1 0 c\u03d1 \u23a4 \u23a6 "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003417_978-3-642-28768-8_14-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003417_978-3-642-28768-8_14-Figure1-1.png",
+        "caption": "Fig. 1 Model of the transmission",
+        "texts": [
+            " We expect that these findings combined with non-stationary load condition are an important problem both from scientific and practical point of view. It is worth to notice that such combination (complexity of gearbox and varying load) can be often met in reality (wind turbines, mining machines, helicopters, etc.) [1,2,7] The model formulation will be first presented incorporating a time varying mesh stiffness with relation to variable load. Numerical simulations and experimental validations will be finally presented and discussed. In this section, a two stage gearbox model is presented (fig. 1). This model will include both local damage and variable load/speed conditions. The model is composed of two pinions and two wheels supported by three bearings, one for the input shaft, the second for the intermediate shaft and the third for the output shaft. The transmission is driven by a motor which transmits torque to the input shaft by a coupling. A load is applied at the end of output shaft. The system has 12 degrees of freedom (DOF) which can be detailed as follows: - Translations of input block (having the mass m1) composed of motor, input, coupling, and pinion1 along x (horizontal) and y (vertical) directions"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003776_j.mechmachtheory.2012.03.003-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003776_j.mechmachtheory.2012.03.003-Figure1-1.png",
+        "caption": "Fig. 1. Circles on the surface of the classical torus (A) and the Bennett linkage whose links overlap with the radii of this torus.",
+        "texts": [
+            "003 Contents lists available at SciVerse ScienceDirect Mechanism and Machine Theory j ourna l homepage: www.e lsev ie r .com/ locate /mechmt mobility of the Bennett mechanism. This mechanism is related to surface, and its mobility is to be calculated in the same manner as the mobility of plane and spherical mechanisms. The mobility calculated on the basis of Chebychev\u2013Gr\u00fcbler\u2013Kutzbach's criterion is equal to unity. The classic torus and its parameters are presented in paper [15]. This is a surface which is formed by rotating a circle with radius s1 around an axis of a circle with radius r1 (Fig. 1A, B). The planes of these circles are perpendicular to one another. The classic torus is also formed by rotating a circle with radius s2 around an axis of a circle with radius r2. These tori will be identical if Villarceau conditions are fulfilled: s1=r2, s1/r1=sin \u03c72. It can be noted that these are also the primary conditions for the Bennett linkage [1,10]: r1=s2,s1=r2,s12 sin 2\u03c71=s2 2 sin 2\u03c72, with \u03c71=\u03c0/2. The Bennett linkage will be formed if the axes of the circles that pass through points Q, R1, R2, and S are revolute pairs which connect links QR1=r1, R1S=s1, QR2=r2, R2S=s2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003888_s00021-012-0105-2-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003888_s00021-012-0105-2-Figure7-1.png",
+        "caption": "Fig. 7. Moving \u03a01 toward \u03a02, for fixed data on the plates. The contact angles \u03b3+ 1 and \u03b3\u2212 1 reappear on I and on V",
+        "texts": [
+            " From this result we obtain immediately Corollary 5.5. If S2 : u(x; a) is a solution of (2.4) that meets \u03a02 in angle \u03b32 = \u03b30 2 and \u03a01 in angle \u03b31 < \u03c0 \u2212 \u03b30 2 , then u (x; a) > u0 2 (x; a) throughout the interval between the plates. Similarly, if \u03b31 > \u03c0 \u2212 \u03b30 2 then u (x; a) < u0 2 (x; a). Any solution is uniquely determined by the angles with which it meets the two plates. We may now complete the proof of the theorem. Decreasing of the separation 2a can be effected by moving the plate \u03a01 to the right (Fig. 7) without disturbing \u03a02. In this way the curve S0+ remains unvaried during the procedure (as it is completely determined by the datum \u03b32 on \u03a02 and by the condition at negative infinity) while the contact angle it forms with \u03a01 increases continuously as \u03a01 is moved toward \u03a02, from \u03b30+ 1 to \u03c0 \u2212 \u03b32. When the trajectory S+ 2 \u2208 Il a2, the crossing point x01 < 1 2 (x1 + x2), and we see directly (Fig. 7) that the angles \u03b30+ 1 and \u03b3+ 1 with which S0+ and S+ 2 meet \u03a01 are such that \u03b30+ 1 < \u03b3+ 1 < \u03c0 \u2212 \u03b32 (5.19) in view of the convexity of S+ 2 . But throughout the procedure \u03b3+ 1 remains unchanged, while \u03b30+ 1 increases continuously to \u03c0 \u2212 \u03b32. Thus there is a unique value a = a\u2217 at which \u03b30+ 1 = \u03b3+ 1 . When this position is attained, the curve S+ 2 meets \u03a01 and \u03a02 in the same angles as does S0+, and hence by the above Corollary must coincide with S0+, which yields zero net force on the plates. This conclusion establishes the initial sentence of the theorem in the case considered",
+            " Hence the corollary now yields that the considered solution u+(x;a) lies above S0+, achieves a positive minimum u+ 0 = u(x0; a) > 0, and thus by (4.1) yields the attracting force \u03c3\u03ba ( u+ 0 )2 between the plates. This completes the proof of the theorem, for the case S2 \u2208 Il a2. When S2 \u2208 Ir a2 analogous reasoning yields the same result, with S0+ replaced by S0\u2212. Here the inequality (5.19) is replaced by \u03b30\u2212 1 > \u03b3\u2212 1 > \u03c0 \u2212 \u03b32. (5.20) A technical change is needed in the reasoning, in the event that S0\u2212 does not extend to \u03a01, as indicated in Fig. 7. In that event, \u03b30\u2212 1 in (5.20) is to be replaced by the inclination \u03c0 achieved by S0\u2212 at the point x4 indicated in the figure. We wish to determine the forces acting on the plates, in terms of the contact angles \u03b31, \u03b32 and the separation distance 2a; we seek general estimates for these forces depending on parameters of the configuration, and notably we seek asymptotic growth or decay estimates as a \u2192 or \u221e. We see from (5.5) that for given materials the problem devolves entirely on determining the angle \u03c80 with which the (unique) solution curve arising from the data \u03b31, \u03b32 intercepts the x-axis",
+            "1) by x (\u03c8) = x2 \u2212 1\u221a 2\u03ba \u03c0 2 \u2212\u03b32\u222b \u03c8 cos \u03c8\u221a sin \u03b31 1 \u2212 cos \u03c8 d\u03c8 u (\u03c8) = \u221a 2 \u03ba (sin \u03b31 1 \u2212 cos \u03c8) (6.5) We find now the intercept height on \u03a02 u2 = \u221a 2 \u03ba (sin \u03b31 1 \u2212 sin \u03b32) (6.6) and for given plate separation 2a = 1\u221a 2\u03ba \u03c0 2 \u2212\u03b32\u222b \u03c81 1 cos \u03c8\u221a cos \u03c81 1 \u2212 cos \u03c8 d\u03c8. (6.7) We may calculate \u03c81 1 from (6.7) as follows: Lemma. Given any 2a > 0 and \u03b32 \u2208 (0, \u03c0/2), there exists a unique \u03c81 1 \u2208 (0, \u03c0/2) for which (6.7) holds. The relations (6.5) then provide the unique solution of (2.4) meeting \u03a02 in angle \u03b32 and \u03a01 in angle \u03b31 1 = \u03c81 1 + \u03c0/2, as indicated in Fig. 7 Proof. We denote the integral appearing in (6.7) by J (\u03c81 1, \u03b32).Setting cos\u03c8 = \u2212s, in the interval \u2212 cos\u03c81 1 < s < \u2212cos((\u03c0/2) \u2212 \u03b32) = \u2212sin\u03b32 (6.8) and setting s1 = \u2212cos\u03c81 1, s2 = \u2212sin\u03b32, we find J (\u03c81 1; \u03b32) = \u2212 s2\u222b s1 s\u221a s \u2212 s1 ds\u221a 1 \u2212 s2 . (6.9) One sees directly that J can be made arbitrarily large by letting \u03c81 1 \u2192 0, also that it becomes arbitrarily close to zero on letting \u03c81 1 increase toward (\u03c0/2)\u2212\u03b32. Thus all positive values can be achieved by adjusting that parameter. We obtain the uniqueness by showing that J is strictly decreasing in s1, and thus also in \u03c81 1: We cannot differentiate (6"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001451_978-0-387-09643-8-Figure7.19-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001451_978-0-387-09643-8-Figure7.19-1.png",
+        "caption": "Figure 7.19: Robot Goes to Destination without S-Net; with noise variance = 10 (adapted from [20]).",
+        "texts": [
+            " The values obtained by each sensor (including on-board sensors and all the SELs) are normally distributed with parameters (\u03bc, \u03c32), in which \u03bc is the ideal temperature value, and \u03c32 is the variance. In fact, each sensor smooth its data value by taking ten samples and returning the average, for both systems with or without the S-Net. From these figures, we can clearly see that when the noise variance is above 10, for the mobile robot that utilizes the S-Net, the result does not change much. But for the mobile robot that uses on-board sensors, it so happens that the robot fails to locate the temperature source correctly (Figure 7.19). The times and distances traveled by the mobile robot depend on this limit; the maximum time allowed for the task is 15 time units. In theory, it might never locate the source. This is because the four on-board sensors are located too close to each other, so the temperatures they report are too noisy to be useful. When noise is added to each sensor, the gradient computed from their values can have large error, which will further change the direction the mobile robot moves. One proposed solution to this problem is to have the mobile robot move to four widely spaced locations and get samples across a greater spatial scale to compute the correct gradient"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000792_aqtr.2008.4588849-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000792_aqtr.2008.4588849-Figure1-1.png",
+        "caption": "Figure 1. The co-ordinate systems and notations for the two-wheeled differential drive mobile robot.",
+        "texts": [
+            " The explanation is that it assures a good balance between large capabilities in locomotion (or tracking possibilities) and mechanical complexity (or associated construction costs) [4]. To characterize the current localization of the mobile robot in its operational space of evolution, we must define first its position and its orientation. The position of the mobile robot on a plane surface is given by the vector ( )yx, , which contains the Cartesian coordinates of its characteristic point P (see Fig. 1). Usually, this characteristic point P is placed in the middle of the common axis of the driven wheels. As we can see in Fig. 1, the orientation (or direction) of the differential mobile robot is given by the angle \u03b8 between the instant linear velocity of the mobile robot v (or the RMX axis) and the local vertical axis. The instant linear velocity of the differential mobile robot v is attached and defined relative to the characteristic point P. As equation (1) denotes, this instant linear velocity is a result of the linear velocities of the left driven wheel Lv and respectively of the right driven wheel Rv . These two drive velocities Lv and Rv are permanently two parallel vectors and, in the same time, they are permanently perpendicular on the common mechanical axis of these two driven wheels",
+            " They give finally the first two state equations (for the linear velocity components of the mobile robot): \u03b8sin 2 \u22c5+= \u2022 RL vvx (5) \u03b8cos 2 \u22c5+= \u2022 RL vvy (6) If we note by RRLL yxyx , , , the Cartesian positions of the driven wheels in the global references attached to the operational space, we can write the next two equations: \u03b8cos\u22c5\u2212=\u2212 ARL lxx (7) \u03b8sin\u22c5=\u2212 ARL lyy (8) and respectively the associate equations: \u03b8\u03b8 sin\u22c5\u22c5=\u2212 \u2022\u2022\u2022 ARL lxx (9) \u03b8\u03b8 cos\u22c5\u22c5=\u2212 \u2022\u2022\u2022 ARL lyy (10) Because the vectors for linear speed of wheels Lv and Rv are orthogonal on the common axis of the driven wheels (see Fig. 1), we can write the third state equation (11), representing the angular velocity of the robot: A RL l vv \u2212= \u2022 \u03b8 (11) The last two state equations denoting the linear accelerations of the two drive wheels are evident: LL av = \u2022 (12) RR av = \u2022 (13) The curvature coefficient (k) associated on a specific trajectory-segment is defined as the inverse ratio of the radius of that trajectory\u2013segment. The equation for the curvature can be obtained because the radius of the trajectory\u2013segment can be writing as a ratio between the linear velocity and the angular velocity of the robot body"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003859_gt2013-95074-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003859_gt2013-95074-Figure1-1.png",
+        "caption": "Figure 1: Coordinate system for force and displacement components for a tilting-pad bearing",
+        "texts": [
+            " This assumption is reasonable for rotors with practically inevitable residual unbalance. Neglecting inertia effects in the oil film force and only considering its dependence from shaft displacement and velocity, leads to the definition of stiffness and damping coefficients. With the oil film force = , shaft displacement = , shaft displacement velocity = , the resulting eight coefficients for a steady-state operating point (index st) are: = ; = ; = , ; = , (1) The directions are defined according to Figure 1. According to all reviewed literature the K-C model is valid to characterize the dynamic properties of fixed-pad bearings in wide ranges, e. g. [10], [11]. Only if fluid inertia forces are of significance the consideration of an added mass term can be required as for example shown by Reinhardt and Lund [12] and Dousti et al. [13]. However, the dynamic behavior of tilting-pad bearings in general needs a more detailed model. First, the elasticity of the pivot influences the properties of the bearing [14]-[16]"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000519_s11071-006-9176-z-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000519_s11071-006-9176-z-Figure1-1.png",
+        "caption": "Fig. 1 Homogeneous disk in motion on an inclined plane",
+        "texts": [
+            " The influence of the coefficient of kinetic fric- tion and the geometrical characteristics (the impact angle, the length and the contact radius of the beam) on the energy dissipated by friction during impact is analyzed. For the double pendulum, using the kine- matic coefficient of restitution, in some cases, can lead to energetically inconsistent results. If the mo- ment of rolling friction is introduced, this problem can be solved for some values of the coefficient of rolling friction. A homogeneous disk in motion on an inclined plane is shown in Fig. 1. The fixed cartesian reference frame x Oyz is chosen with the origin at O . The angle between the axis Ox and the horizontal is \u03b1. The contact point between the disk and the plane is B. The disk has the mass m, the radius r , and the center of mass at C . The gravitational acceleration is g. 2.1 Pure rolling (no sliding) The forces that act on the disk are the gravitational force G at point C , the normal reaction force N of the plane and the friction force F f at the contact point B. The rolling friction is considered negligible"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003045_s11044-012-9326-7-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003045_s11044-012-9326-7-Figure2-1.png",
+        "caption": "Fig. 2 A model of the planar telescopic-legged rimless wheel",
+        "texts": [
+            " Second, we extend the method to a planar telescopic-legged biped robot with flat feet incorporating a brake spring and numerically investigate the gait descriptors, especially the behavior of ZMP. Through numerical simulations, we discuss the anterior-posterior asymmetry of human foot from the ZMP point of view. This paper is organized as follows. Section 2 investigates the validity of the method using a simple rimless wheel model with telescopic legs. Section 3 extends the method to a planar telescopic-legged biped model with flat feet and ankle brake. Finally, Sect. 4 concludes this paper and describes the future research directions. 2.1 Modeling of telescopic-legged rimless wheel Figure 2 shows the model of a planar telescopic-legged rimless wheel (TRW). We call the leg frame on the floor \u201cstance leg,\u201d and assume that only the stance leg is actuated and other legs are kept at l0 [m]. We also assume that this model has a point mass, M [kg], concentrated on the central position, and the inertia moment about the CoM can be neglected. In this paper, we set \u03b1 = \u03c0/4 [rad], i.e. eight-legged, and l0 = 1.0 [m]. Figure 3 shows the essential part of the TRW which determines the stance-phase motion"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure1.4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure1.4-1.png",
+        "caption": "Figure 1.4 Mechanical linkage summer/comparator",
+        "texts": [
+            " Mechanical Linkage Summing Mechanical linkage summing is an extremely simple form of summing device that is used extensively in many aircraft flight control systems in service today. Movements of the pilot\u2019s control column are typically translated via cables and pulleys or push\u2013pull rods to mechanical inputs to servo control actuators at the control surface. These actuators provide the necessary muscle to overcome the aerodynamic forces associated with high speed flight. The pilot\u2019s input is compared with the control surface position (servo actuator output) in a mechanical summing linkage arrangement similar to that shown in Figure 1.4. In a similar manner inputs from an autopilot actuator can be summed with the pilot\u2019s command to provide an auto-stabilization function. The Speed Governor The speed governor goes back more than 200 years to James Watt who invented the \u2018flyball governor\u2019 as a mechanism to control the speed of a steam engine that did not require human intervention. Derivatives of 8 Developing the Foundation this device are used throughout industries using rotating machinery. Figure 1.5 shows such a device in schematic and block diagram form"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002017_10426914.2010.496126-Figure18-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002017_10426914.2010.496126-Figure18-1.png",
+        "caption": "Figure 18.\u2014Plastic strain contour in the Y -axis of the turbine component after cold isostatic pressing.",
+        "texts": [
+            " There is D ow nl oa de d by [ U ni ve rs ity o f T as m an ia ] at 1 7: 01 0 1 Se pt em be r 20 14 Figure 19.\u2014von Mises stress contour of the turbine component after cold isostatic pressing. a general agreement between the experimental results and the simulation results. The max error is\u22127.27%, which may because the diameter of a small circle is so small that the error of measurement was large and the foot of the small circle had some deformation and distortion. The plastic strain in the X-axis (as shown in Fig. 17) is smaller than that in the Y -axis (as shown in Fig. 18). So the shrinkage of the heights is larger than that of the diameters. The von Mises stress contour of the turbine component (as shown in Fig. 19) is mostly uniform on the whole part, which implies that the plastic deformation is mostly uniform. This shows that the part has only uniform shrinkage of volume without distortion, which is the same as the cylindrical specimen. A combined method of SLS and CIP was applied to manufacture the metal parts. The results show that the method is feasible and effective"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000846_peds.2007.4487880-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000846_peds.2007.4487880-Figure1-1.png",
+        "caption": "Fig. 1. Stator and rotor flux linkages in different reference frames",
+        "texts": [
+            " The four quadrant and field weakening operation of the proposed drive is also studied and results show that the ripples in flux and torque are greatly reduced both below and above base speed with improved estimation ofwinding currents. II. OPERATING PRINCIPLE OF DIRECT TORQUE CONTROL In direct torque controlled PMSM drive, while the amplitude of the stator flux linkage is kept invariant, the electromagnetic torque of a PMSM is directly proportional to the sine value of electrical angle between the stator flux linkage and rotor flux linkage vector. The stator flux linkage (D-Q, i.e. stator frame) and rotor flux linkage (d-q, i.e. rotor 1-4244-0645-5/07/$20.00\u00a92007 IEEE 1354 frame) are represented in Fig. 1 where ds-qs is synchronous rotating frame. Therefore the precise torque control can be achieved by controlling the instantaneous speed of the stator flux linkage with its amplitude keeping constant below base speed and flux weakening above base speed. It is fulfilled through the application of proper selecting the voltage space vector generated by an inverter and therefore fast torque response can be obtained by increasing the rotating speed of the stator flux linkage as fast as possible. III. DSVM-DTC SYSTEM FOR PMSM A",
+            " The phase currents are computed using inverse Park's transformation as ia = d CosOr - iq SinOr (13) ib=idCOS(Or - 21/3) - iqSinf(Or - 21/3) (14) ic 1dCos (Or - 4 qt3)iqSif(Or - 42t3) (15) where, 0, is the position angle of the rotor. Te T. lET-- tk T3/3 T3/3 T3/3 Time (a) tk TS/3 TS/3 T,/3 Time (b) Fig. 5. Comparison of torque waveform when dT,=-1 between (a) classical DSVM-DTC and (b) improved DSVM-DTC V. CURRENT SENSORLESS CONTROL OF PMSM DRIVE A. Current Estimation Rotor reference d-q currents are estimated using motor model in discrete equation based on eq. (16)-(17) in winding current estimation block with sampling time Ts which is feedback to torque and flux estimator block as shown in Fig. 1. Estimated rotor reference d- and q- axes currents (idest, iqest) are calculated as follows i 1jm + 1) =id(m)+ T Vqqsmdest (m+1)idest (m)+Ld {vd (m) +coe (m)Lqi e t(m) -Rsi dest(m)} (16) T qest (m + 1) iqest(m)+L- {vq(m)- \u00b0Oe(m)Pf- \u00b00e (m)Ldidest (m) - Rsiqest (m)} (17) where voltages vd(m) and vq(m) are d- and q- axis components of the stator voltage supplied to the motor at mth sample instant respectively. B. Stator Voltage Estimation The applied three phase stator winding voltages are reconstructed using a voltage sensor mounted at the dclink terminal of the VSI and the applied switching pulses to the inverter legs"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003498_1.4024212-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003498_1.4024212-Figure5-1.png",
+        "caption": "Fig. 5 Stick diagram and multibody model of the experimental transmission",
+        "texts": [
+            " Comparisons showed that the deviations between analytically and numerically determined forces are small. So the presented approach is feasible. If the forces are identical, the determined state variables would also be the same with a completely numerical time integration. Simulations with a simple experimental transmission were done, which was also available as hardware so that validation measurements could be carried out. The stick diagram and a 3D view of the transmission model are shown in Fig. 5. It consists of two shafts with one gearwheel per shaft. The gearwheel on the input shaft is fixed by a shaft-hub joint and the one on the output shaft represents an idler gear which is fixed by a synchronizer. Both shafts are mounted by rolling bearings in the aluminum housing which contains oil to lubricate and cool the transmission. The gear ratio, the helix angle, the gear backlash, and the axial clearance of the idler gear can be changed to investigate the influence of these parameters. For the following simulations, a gear pair with a ratio of 37/58, a helix angle of 20 deg, a tooth width of 20 mm, and a module of 2 was used"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003014_iros.2011.6094585-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003014_iros.2011.6094585-Figure5-1.png",
+        "caption": "Fig. 5. Kinematic model of the ASOC-driven omnidirectional mobile robot.",
+        "texts": [
+            " The angle of rotation of the pivot and roll axes are measured by potentiometers. produce planar translational velocities at a point along its pivot axis by independently controlling each wheel\u2019s velocity. A control method is presented in Section III-B. Each ASOC module is a self-sustained robotic system, comprised of a power supply, actuators, microcontroller (PIC), wireless device, and motor driver. Each module performs simple tasks assigned by the supervisory computer on the body, and executes local feedback control. A. Kinematic Model Fig. 5 illustrates a kinematic model of the ASOC-driven omnidirectional mobile robot. The coordinate frame for the main body \u03a3b is fixed on the body centroid and defined as a right-hand frame, depicting the longitudinal direction as x. The coordinate frame for each ASOC module \u03a3i (i=1\u20264) is defined such that the z axis is aligned to the pivot shaft and fixed at a point along its pivot axis. (\u03a3i does not rotate along with the ASOC rotation around its pivot axis.) D and \u03bei locate each ASOC module with regard to the main body, and r is the wheel radius"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003086_pime_conf_1966_181_311_02-Figure8.3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003086_pime_conf_1966_181_311_02-Figure8.3-1.png",
+        "caption": "Fig. 8.3. Cavitation boundary",
+        "texts": [
+            " Since the relaxation continued beyond the cavitation boundary, the solution would exhibit negative values of q5 4\u2018\u201d\u2019\u201d = D-l[B-(L+ u)+(k\u2019] Prac Instn Mech Engrs 1966-67 in this region, and the boundary condition would not be satisfied as shown in Fig. 8 . 3 ~ . Therefore, a mechanism was included in the programme to find this boundary automatically. After each iteration each negative value of + was set to zero. This had the effect of raising the level of C) in this region, so that in the next iteration the pressures at the points surrounding the one being calculated increased its value slightly until finally no negative pressures were obtained and the boundary condition was satisfied, as shown in Fig. 8.3. It is appreciated that the Reynolds equation gives a solution which would follow the dotted line in Fig. 8.3b if permitted to do so. However, so long as the distance from the cavitation line to the last mesh line is not too great and the mesh is fine in this region, the effect of holding pressures beyond the cavitation line to zero will be small. Two procedures were written to integrate a function over a set of evenly and unevenly spaced points. These were used to calculate the force components from the pressure distribution. The important parameter produced by the computer programme is the lubricant pressure"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003777_irsec.2013.6529676-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003777_irsec.2013.6529676-Figure7-1.png",
+        "caption": "Figure 7. WGS topology and control",
+        "texts": [
+            "1 By maintaining the power coefficient at its maximum, the WGS can produce maximum power for TSR optimum. The optimum torque is then expressed by (13) Where CoPt = K \u00a3lopt 2 _ 1 Rb3 K - - pSCpmax --3 2 Aopt (13) (14) \ufffd 1,500 500 IdOO Figure 6. Power characteristics network According to power balance, we establish the power available at the PMSG stator in accordance to the expression (16) dO PPMSG = Pmec - nu dt + fD.) (16) The PMSG is directly connected to a three phase diode uncontrolled rectifier as shown in figure 7. The average output voltage of the rectifier is obtained as follows in (17). In the figure 8 we present a simulation result of the rectifier output associated to the PMSG driven in different wind speed If we neglect the rectifier looses, the electrical output power from the PMSG can be represented (18) For providing electric power to domestic applications in conformity with the Moroccan Low voltage grid (220V 150Hz) , the DC output voltage of the rectifier has to be inverted in AC form by a single phase double bridge VSI ",
+            " The extra power produced can be stored in the chemical system and used when the wind speed is less than the stall speed. However, the use of an elevator transformer makes the weight of the platform even greater regarding to the weight of the batteries. In addition, the cycle life of the batteries is limited and change increases the cost. The objective of this work is to modify the topology of the platform in order to eliminate the chemical storage and all the power produced will be transmitted to the load or may be injected to a low voltage grid. The new topology is presented in figure7. We intercalate a DC Boost Chopper (BC) between the rectifier and the VSI. Its role is to control the rectifier output voltage and adjust it to the VSI. It consist of a power IOBT with its anti-parallel diode and two basic components an inductor and a capacitor. By switching the BC, the power Pin which is the same as PPMSG is first stored in inductor when the BC is Closed ON and then transferred to the VSI via the capacitor when BC is OFF. The BC output voltage is expressed in (20) v.: = _T_ v:",
+            " CONTROL (20) Since the wind power fluctuates with wind velocity, the PMSG output voltage, frequency and the dc link voltage Vin of the rectifier side vary continuously. The boost chopper operating in continuous conduction mode (CCM) controls the DC voltage by controlling the switch duty cycle a to obtain a constant voltage. The PMSG rotor speed is then controlled by regulating the inductance current iL of the boost converter. This control scheme aims to control the shaft speed of the PMSG driven wind turbine in order to keep the TSR at the optimal point and extract Maximum power for all wind speeds. The DC boost converter and its control is done in figure 7 and detailed in figure II. Vc ref figure 11. Block diagram of booth current and voltage control loop This control is composed by two nested loops with PI controller. The main loop of voltage provides the reference for the interior current loop. The voltage reference is dictated by the output of the VSI in accordance with the grid. A simplified equivalent small-signal model of the BC in continuous conduction mode is given by [19]. Gi and Gv represent respectively the transfer function useful to tune the current and voltage loop control are given in (21) G"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003091_aim.2010.5695776-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003091_aim.2010.5695776-Figure1-1.png",
+        "caption": "Fig. 1. Model of the passive dynamic walker with flat segmented feet and compliant ankle joints.",
+        "texts": [
+            " The proposed segmented foot model is compared with the rigid foot model to reveal the advantages of adding toe joint to flat foot. This paper is organized as follows. Section II describes the model in detail. In Section III, we show the experimental results. We conclude in Section IV. To obtain further understanding of real human walking, we propose a passive dynamic bipedal walking model that is more close to human beings. We add compliant ankle joints and flat segmented feet with compliant toe joints to the model. As shown in Fig. 1, the two-dimensional model consists of two rigid legs interconnected individually through a hinge. Each leg contains segmented foot. The mass of the walker is divided into several point masses: hip mass, leg masses, masses of foot without toe, toe masses. Each point mass is placed at the center of corresponding stick. Torsional springs are mounted on both ankle joints and toe joints to represents joint stiffness. To simplify the motion, we have several assumptions, including that legs suffering no flexible deformation, hip joint with no damping or friction, the friction between walker and ground is enough, thus the 978-1-4244-8030-2/10/$26",
+            " And foot-scuffing at mid-stance is neglected since the model has no knee joints. We suppose that the x-axis is along the slope while the y-axis is orthogonal to the slope upwards. The configuration of the walker is defined by the coordinates of the point mass on hip joint and six angles (swing angles between vertical coordinates and each leg, foot angles between horizontal coordinates and each foot, toe angles between horizontal coordinates and each toe), which can be arranged in a generalized vector q = (xh, yh, \u03b11, \u03b12, \u03b11f , \u03b12f , \u03b11t, \u03b12t) T (see Fig. 1). The positive direction of all the angles are counter-clockwise. The segmented foot structure used in this paper is shown in Fig. 2. The foot mass is distributed at two point masses: one at the center of toe, and the other at the center of the rest part of the foot (ms and mf \u2212ms in Fig. 2). We define foot ratio as the ratio of distance between heel and ankle joint to distance between ankle joint and toe tip, namely a/b in Fig. 2, but not a/c, to make it convenient to compare the proposed model with rigid foot model"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001719_s11044-009-9165-3-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001719_s11044-009-9165-3-Figure3-1.png",
+        "caption": "Fig. 3 The coordinate frames for contact kinematics",
+        "texts": [
+            " Therefore, a surface point c can be represented by a parametric form pbc = pbc(u, v). Here, pbc is the vector from the origin of frame b to c, which is represented in frame b. We can also define a surface frame at the point c with its transformation matrix denoted as Tbc = (Rbc,pbc), where Rbc = [xbc, ybc, zbc] (1) with xbc = \u2202pbc \u2202u \u2016 \u2202pbc \u2202u \u2016 , zbc = \u2202pbc \u2202u \u00d7 \u2202pbc \u2202v \u2016 \u2202pbc \u2202u \u00d7 \u2202pbc \u2202v \u2016 , (2) ybc = zbc \u00d7 xbc. For a contact of two bodies f and g, we attach the coordinate frames as shown in Fig. 3. Here, cf and cg are the contact frames which are always defined on the contact (or the nearest) points of f and g. lf and lg are local frames. They are fixed on the contact points of body f and g, and coincide with lf and lg, respectively, at the time of interest. f and g also refer to the name of the body frames, while w represents the inertia reference frame. The relative motion of the two bodies has six degrees of freedom in general and can be expressed by the four parameters of the surfaces (uf , vf , ug, vg), the distance of the nearest points d , and the angle of contact \u03c8\u2014the angle between the x-axes of cf and cg "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-107-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003218_b978-0-12-386045-3.00003-9-Figure3-107-1.png",
+        "caption": "Figure 3-107: Pad carrier ring damaged by axial shaft vibration.",
+        "texts": [
+            " Machinery Component Failure Analysis 171 cycling through an excessive temperature range. In all three cases, it would be appropriate to investigate and, if possible, reduce maximum operating temperatures. Fit new or reconditioned pads. Axial vibration can cause damage and fatigue of tilting-pad pivots. Both pivot and carrier ring may suffer damage by indentation or fretting. In some cases tiny hemispherical cavities may be produced. Damage may occur due to axial vibration imposed upon the journal, or may be caused by the thrust collar face running out of true. In Figure 3-107, the pad carrier ring shows damage due to axial shaft vibrations. Similarly, Figure 3-108 depicts a pad pivot with hemispherical cavities caused by pivot fretting due to vibration. Here, it will be necessary to recondition the carrier ring and fit new or reconditioned pads. Investigate and eliminate the cause of axial vibration. Someone once said that \u201cGears wear out until they wear in, and then they wear forever.\u201d The American Gear Manufacturers Association (AGMA) describes this mechanism more clearly as follows: \u201cIt is the usual experience with a set of gears on a gear unit "
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002587_j.simpat.2010.04.002-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002587_j.simpat.2010.04.002-Figure5-1.png",
+        "caption": "Fig. 5. System configuration of rotor-injection.",
+        "texts": [
+            " The electromagnetic torque will increase until it equals the applied mechanical torque. The stator voltages observed in the synchronous reference frame have a high frequency in the beginning and become constants finally. This is due to the fact that the abc variables have a frequency variation from 15 Hz to 60 Hz. Observed from the synchronous rotating reference frame, the frequency varies from 45 Hz to 0 Hz. The circuit diagram of a DFIG with rotor injection and an isolated load on the stator is shown in Fig. 5. The injected AC voltage to the rotor usually comes from a DC/AC bridge converter shown in Fig. 6. Pulse Width Modulation (PWM) is widely used to control the frequency and amplitude of the output of the rotor-side converter. To simplify the control circuit and reduce the switching losses, a six-step switching technique which results a quasi-sine voltage is sometimes applied to the rotor circuit. The quasi-sine waveform does not need the sine and triangular waves used in a conventional sine PWM implementation and the frequency variation is also equally simple"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001699_s1068798x10100072-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001699_s1068798x10100072-Figure2-1.png",
+        "caption": "Fig. 2. Components of transmission with intermediate rollers and a free iron ring: (1) cam; (2) roller (ball or cyl inder); (3) ring (separator); (4) crown.",
+        "texts": [
+            " This theorem establishes a relation between the centers of the centroid and the centers of radii of curvature of the gear profiles in engagement. With known gear centroids and the rela tions between them and the coordinates of the cen troid\u2019s centers of curvature [4], we may derive expres sions for radii of curvature of the cycloid gear profiles in terms of the initial parameters of the transmission with intermediate rollers and a free iron ring. We now consider the cross section of a transmission with intermediate rollers and a free iron ring (Fig. 2). Note that the roller is between two cycloid profiles: the cam and the crown. In Fig. 2, P is the engagement DOI: 10.3103/S1068798X10100072 1002 RUSSIAN ENGINEERING RESEARCH Vol. 30 No. 10 2010 EFREMENKOV, AN I KAN pole; O1 is the center of the cam; O2 is the center of the circle defined by the centers of the rollers and the ring; O3 is the center of the crown; rw1, rw2, rw3 are radii of the cam centroid, the ring, and the crown, respectively. The initial parameters for a transmission with inter mediate rollers and a free iron ring are as follows [4]: rw2, the radius of the generating circle; Z2, the number of rollers; \u03c7, the bias; and rr, the roller radius. It is evident from Fig. 2 that the cycloid surfaces of the cam and the crown have a transitional profile: con cave in the recesses and convex in the projections. Thus, the roller traveling around the profile from a profile valley to a projection passes from a concave section of the profile to a convex section. The roller corresponds to a cylinder or sphere and has a convex profile with constant radius of curvature. Structural analysis suggests that this transmission may be regarded as two coupled transmissions [4, 5]: engage ment of the cam with the rollers and ring; and engage ment of the rollers and ring with the crown"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001770_acc.2010.5531275-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001770_acc.2010.5531275-Figure2-1.png",
+        "caption": "Fig. 2. Illustration of the coordinations in the earth frame (inertial frame) {E}, the ship body-fixed frame {B} and the Serret-Frenet frame {SF}.",
+        "texts": [
+            " In Section III, the MPC algorithm is developed to address the path following problem with roll constraints using rudder and propeller. The simulation results are presented in Section IV together with discussions on the tuning of key controller parameters, followed by the conclusions in Section V. SURFACE VESSEL MODEL In the open literature, the path following problem is often simplified into a regulation problem by adopting proper path following error dynamics ( [9], [15]\u2013[18]). For this approach, the Serret-Frenet frame ( [19], [20]) is often adopted to derive the error dynamics. Fig. 2 shows the definitions of the errors used for path following control. The origin of the frame {SF} is located at the closest point on the path curve C from the origin of the frame {B}. The error dynamics based on the Serret-Frenet equations are introduced in [18], given as: \u02d9\u0304\u03c8 = \u03c8\u0307 \u2212 \u03c8\u0307SF = \u03ba 1\u2212 e\u03ba (usin\u03c8\u0304 \u2212 vcos\u03c8\u0304)+ r, (1) e\u0307 = usin\u03c8\u0304 + vcos\u03c8\u0304, (2) where e, defined as the distance between the origins of {SF} and {B}, and \u03c8\u0304 := \u03c8 \u2212\u03c8SF , are referred to as the crosstrack error and heading error respectively, u, v, r are the surge, sway and yaw velocity respectively. \u03c8 is the heading angle of the vessel and \u03c8SF is the path tangential direction as shown in Fig. 2 [18], \u03ba is the curvature of the given path. The control objective of the path following problem is to drive e and \u03c8\u0304 to zero. The path for surface vessels to follow in the open sea is often a straight line or a way-point path, which consists of piecewise straight lines. In these cases, the curvature \u03ba is zero, therefore the heading error dynamics (1) could be simplified as: \u02d9\u0304\u03c8 = r. (3) Marine surface vessels have 6 degrees of freedom. For maneuvering of surface vessels, normally 3 DoFs are of interest, namely the surge, sway and yaw"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003689_20121023-3-fr-4025.00027-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003689_20121023-3-fr-4025.00027-Figure1-1.png",
+        "caption": "Fig. 1. Drive line of an electrically driven vehicle.",
+        "texts": [
+            " In Section 3, observability measures for two different sensor configurations are analyzed. In Section 4, a second order sliding mode observer for linear systems with unknown input is described. Based on the observer, a generalized second order sliding mode controller is applied in Section 5. In 978-3-902823-16-8/12/$20.00 \u00a9 2012 IFAC 79 10.3182/20121023-3-FR-4025.00027 Section 6, the controller is verified with a MBS model of an electrically driven rear axle of a prototype hybrid vehicle. The drive line to be modeled is shown in Fig. 1. The backlashes of the entire drive line are lumped. In order to model the drive line as a two-mass system, the wheels and the drive line elasticities are lumped as well. The elasticity of the drive line, which mainly consists of the elasticities of the side shafts, is represented by a spring with stiffness ks = ks,l + ks,r. The differential equations of the drive line model can be summarized as \u03b8\u0307 = \u03c9m kg \u2212 \u03c9l, Jm \u03c9\u0307m = \u2212 Ts(\u03b8) kg \u2212 dm \u03c9m + Tm, Jl \u03c9\u0307l = Ts(\u03b8) \u2212 dl \u03c9l \u2212 Tl, (1) where Jm and Jl respectively denote the motor and load inertia"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002158_j.advengsoft.2009.12.019-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002158_j.advengsoft.2009.12.019-Figure5-1.png",
+        "caption": "Fig. 5. Disk parameterization methods.",
+        "texts": [
+            " The isotropic Inconel718 has a density nearly four times greater than the anisotropic Epoxy\u2013 Fiberglass composite material. Careful investigation of Figs. 2 and 3 shows that the maximum tangential stress is much higher in the Inconel718 case, which is consistent with the expected trends. When optimizing a component, it is necessary to have the geometry defined using a finite number of modifiable design parameters. In this investigation five geometry parameterization methods were used. Representative disks defined by four of the methods are shown in Fig. 5. The main purpose of these parameterization methods is to define the live weight shown in Fig. 5. Permutations of the Ring, Web, and Hyperbolic disk definition methods have been previously used in other disk design programs [1,2]. TAxi Disk also uses these methods along with the Continuous Slope (CS) parameterization, which was developed specifically for this investigation. A much more detailed description of these geometry definitions is included in the T-Axi Disk documentation. The Ring parameterization is by far the simplest disk definition method. Only one parameter; the bore radius, is needed to define the live weight of this type of disk"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001109_s11740-008-0080-x-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001109_s11740-008-0080-x-Figure3-1.png",
+        "caption": "Fig. 3 Four different modulation strategies: v linear scan velocity, r modulation radius",
+        "texts": [
+            " In the second set of experiments the modulation strategy was varied. Beside the circular modulation four other symmetric strategies were performed. These cross sections were also investigated regarding the homogeneity, shape of the track, defects and especially the penetration depth of the track. The objective of the second experiment was to determine possible influences on the distribution of the filler material by the modulation. In addition the shape of the track is supposed to be influenced. The four modulation strategies are shown in Fig. 3. The \u2018\u2018linear perpendicular\u2019\u2019 modulation is a simple linear one-dimensional modulation between the left and right border of the track. The \u2018\u2018eight\u2019\u2019 modulations consist of two circles. In the \u2018\u2018eight forward\u2019\u2019 modulation the local motion velocity of the laser spot is larger at the border in comparison to the middle. In the \u2018\u2018eight backward\u2019\u2019 modulation the local scan velocity of the laser spot is larger in the middle. The \u2018\u2018circle overlaid\u2019\u2019 modulation is a circle modulation overlaid with a second faster circle modulation"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001559_acemp.2007.4510568-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001559_acemp.2007.4510568-Figure8-1.png",
+        "caption": "Fig. 8. Homopolar path of the permanent magnets flux",
+        "texts": [
+            " However, this stays anought low (4,2%) and allows the validation of the modelisation of the 2D flux path of the homopolar hybrid excitation synchronous machines. B. 3D flux paths In hybrid excitation synchronous machines, the flux paths are not only 2D paths but 3D paths exists also. Those paths can be classed in three categories. The first categorie correspond to paths that cross only once the air gap, that is, they pass through only one pole (homopolar paths). They lead to a non-zero mean value of the flux (Fig. 8). The second category is the one of paths that cross the air gap twice, once through each poles (bipolar paths). The third categorie contains all leackge paths. One of them is illustrated in the Fig. 9. It\u2019s a possibel path for the permanent magnets leackage flux. Indeed, a part of the permanent magnets flux doesn\u2019t cross the actif air gaps and stay localised only in the rotor [3]. The final magnetic equivalent circuit of the homopolar hybrid excitation synchronous machine is obtained by using the develloped analytical model for 2D paths and integrating the 3D paths defined above (Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001321_j.mechmachtheory.2007.08.001-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001321_j.mechmachtheory.2007.08.001-Figure2-1.png",
+        "caption": "Fig. 2. Delta robot.",
+        "texts": [
+            " Similarly, for parallel mechanism that has two chains, i.e., single closed loop mechanism, the number of row of its coefficient matrix is always 6, so if the total number of joints is more than 6, its mobility must be 1 at least. Therefore, one important conclusion can be obtained: If one mechanism is not complicated parallel mechanisms, and the total number of joints is more than 6 (a- 1), its mobility must be 1 at least, where, a is the number of chains. For each chain of the parallel mechanism (Delta robot) in Fig. 2, all axes of revolute joints are parallel with each other. Chains are not simple serial mechanisms, so it belongs to complicated parallel mechanisms. For complicated parallel mechanisms, complicated parallel chains should be changed into equivalent serial chains above all. One parallelogram mechanism is in every chain of Delta robot. Spherical joints can be simplified to revolute joints in view of local degrees-of-freedom, whose axes are vertical to the plane of parallelogram mechanism (Fig. 3)",
+            " If the coordinates of points A, B, C and D are ( 1, 2,0), ( 1,2,0), (1, 2,0) and (1, 2,0), respectively, and reference point is B, then Jacobian matrixes JAB and JDC are: J AB \u00bc 0 0 0 0 1 1 4 0 0 0 0 0 2 666666664 3 777777775 \u00f010\u00de J DC \u00bc 0 0 0 0 1 1 4 0 2 2 0 0 2 666666664 3 777777775 \u00f011\u00de The coefficient matrix is: J \u00bc J AB J DC\u00bd \u00f012\u00de The singular value decomposition of coefficient matrix is: V \u00bc 5:87 0 0 0 0 2:95 0 0 0 0 0:93 0 0 0 0 0 0 0 0 0 0 0 0 0 2 666666664 3 777777775 \u00f013\u00de Eq. (13) means that degree-of-freedom of parallelogram mechanism is 1. Considering the character of parallelogram mechanism, there is _hA \u00bc _hB _hD \u00bc _hC ( \u00f014\u00de It is clear that the unique degree-of-freedom is translational along axis x because 0 0 0 0 1 1 4 0 0 0 0 0 2 666666664 3 777777775 _hA _hB \" # \u00bc 0 0 0 0 1 1 4 0 2 2 0 0 2 666666664 3 777777775 _hD _hC \" # \u00f015\u00de So the parallelogram mechanism can be changed into a revolute joint whose axis is parallel with axis x. As Fig. 2, three chains are symmetrical and corresponding points are located at the same circle. Parallelogram mechanism is simplified as a revolute joint whose axis is vertical to the plane of parallelogram mechanism and center point is coplanar with moving platform. The coordinates of point O, O 0, O00, A, A 0, A00, B, B 0, B00 and reference point are (1, 0,1), ( 0.5,0.866, 1), ( 0.5, 0.866,1), (2,0,0.75), ( 1,1.732, 0.75), ( 1, 1.732, 0.75), (1.25,0,0), ( 0.625, 1.0825,0), ( 0.625, 1.0825,0) and (0,0,0), respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000544_1.2991175-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000544_1.2991175-Figure1-1.png",
+        "caption": "Fig. 1 Connecting rod CAD and mesh",
+        "texts": [],
+        "surrounding_texts": [
+            "1 i t\nt s v\nf t\nt s o c c p\nt p s g n\nq a\nT A C\nJ\nDownloaded Fr\nArthur Francisco e-mail: arthur.francisco@univ-poitiers.fr\nAurelian Fatu\nDominique Bonneau\nLaboratoire de M\u00e9canique des Solides, Universit\u00e9 de Poitiers,\nUMR CNRS 6610, 4 Avenue de Varsovie,\n16021 Angoul\u00eame Cedex, France\nUsing Design of Experiments to Analyze the Connecting Rod Big-End Bearing Behavior Reducing the frictional loss in internal combustion engines (ICE) represents a challenge, in which all car manufacturers are involved. This concern has two origins. The first one is the fuel cost, which increases over the years. The second is strongly linked to ecology: people feel more and more concerned by the greenhouse effect, partly resulting from fuel consumption. Many projects involving several laboratories and lead by car manufacturers have this particular point as main subject, with the goal to reduce the ICE fuel consumption by decreasing the friction power loss. This aim can be partly achieved with a better knowledge of the connecting rod big-end bearing functioning. A lot of theoretical and experimental studies have been carried out, resulting in efficient models for numerical simulations, but at the time, no known ambitious parametric study has been planned, to determine the most influent parameters and to quantify their effects on power loss. The present work is a first step to bridge the gap between the potential of recent numerical simulations and the need for a better understanding of the connecting rod big-end bearing functioning. To plan the numerical simulations, it will be taken advantage of design of experiment techniques, which provide an efficient way of preparing the series of experiments with a minimum of runs. Thus, these techniques are illustrated through the variable combination run, test results generated, and interpretations made to identify the dominate factors impacting the responses of interest. DOI: 10.1115/1.2991175\nKeywords: TEHD lubrication, connecting rod big-end bearing\nIntroduction\nTo reduce exhaust emissions and fuel consumption, technology s being developed either to modify conventional automobile inernal combustion engine ICE or to provide them alternatives.\nAs concerns the first item and in a nowadays competitive conext, engineering centers have to find the parameters for which ome operating values are optimal, from a consumption point of iew.\nThis work will focus on a particular part of the ICE: The transormation of the piston translation motion into the crankshaft roation motion and especially the connecting rod big-end bearing.\nEven if this connection appears quite simple and well conrolled, it still remains the place of important power loss. Can this ystem be really optimized? The problem arises with the number f parameters to deal with: Keeping the connecting rod geometry onstant, the oil properties, the bearing geometry, the oil supply haracteristics, etc., still remain to optimize, i.e., at least ten arameters.\nThus, it becomes unreasonable to find analytically which of hem have to be changed and how to decrease the dissipated ower loss, for example, in the same time the temperature. In the ame way, decreasing the bearing length, for downsizing investiations, can affect dramatically the operating temperature with evertheless a substantial friction loss reduction.\nThe aim of the present work is to numerically and roughly uantify the impact of each variable on power loss, temperature, nd flow rate global responses but not to optimize them this\nContributed by the Tribology Division of ASME for publication in the JOURNAL OF RIBOLOGY. Manuscript received September 24, 2007; final manuscript received ugust 25, 2008; published online December 2, 2008. Assoc. Editor: Shuangbiao Jordan Liu. Paper presented at the ASME /STLE 2007 International Joint Tribology\nonference TRIB2007 , San Diego, CA, October 22\u201324, 2007.\nournal of Tribology Copyright \u00a9 20\nom: http://tribology.asmedigitalcollection.asme.org/ on 01/28/2016 Terms\nambitious task is reserved for further works . This way, it becomes possible to find a reasonable compromise for the kind of situation described.\nThe software ACCEL, developed in our laboratory, will be described through the models used and the input parameters needed for a thermoelastohydrodynamic TEHD calculation. The ten variables chosen to study the global responses of the connecting rod bearing will only have two possible levels, a low and a high level. Hence, with no special strategy, it costs 210 calculations in order to explore all the possible cases, but this is not realistic.\nThat is the reason why design of experiment DOE techniques will be used. It will make possible achieving the primary goal what are the most influent parameters in the bearing functioning , with acceptable calculation times. These techniques will be explained, as well as the particular DOE used to reduce the number of runs.\nBecause of the approximations made due to the reduction in the number of runs , the results have to be analyzed carefully. The DOE chosen leads to an approximated analytical model, for which the errors introduced must be evaluated. Thus, the principal effects of the parameters, on the bearing behavior, will be described within a confidence interval resulting from the introduced approximations.\nFinally, we will discuss other physicals values, such as film thickness and pressure, whose levels must also be maintained while seeking to minimize other parameters such as flow rate, power loss, and temperature.\n2 Paper Background During the past two decades, engine elements have been designed smaller and smaller while increasing up to 30% the power output 1 . This result has been partly obtained, thanks to nearly optimal connecting rod designs. The knowledge required for this particular task implies a good understanding of the connecting rod\nJANUARY 2009, Vol. 131 / 011101-109 by ASME\nof Use: http://www.asme.org/about-asme/terms-of-use",
+            "t l i\ne t d t t c e a t a A\nn t e o\nc f m e t p d F m e c fi m n\nr b t t a c\nm U p\nn a m\nc t t j t\nm\n0\nDownloaded Fr\nhermoelastohydrodynamic behavior. That is the reason why the atter has been the subject of many theoretical and experimental nvestigations since the early 1980s.\nFantino et al. 2\u20134 studied the effect of the deformation of an lastic automotive connecting rod on the oil film characteristics in he big-end bearing. Consistent elastic deformations and pressure istributions were obtained by iterative methods, using plane elasicity relations and a rigid shaft. Finite differences were used for he EHD model with elasticity terms coming from finite element alculi on the housing. Negative pressures were set to zero whenver they occurred. It is interesting to point out that there was yet wish to quantify the effects of some particular parameters, say, he viscosity, the piezoviscosity coefficient, and the radial clearnce, on the bearing behavior through its minimum film thickness. n empirical dimensional equation was then proposed. Goenka and Oh 5,6 presented theoretical work on EHD conecting rod big-end bearing problems. The Newton\u2013Raphson echnique was used in conjunction with Murty\u2019s algorithm. Finite lement method FEM was used to analyze the EHD lubrication f a journal bearing under dynamic loading.\nIn Ref. 7 , the hydrodynamic of a connecting rod bearing was alculated considering not only the effects of bearing elastic deormation but also those of bearing shape. The shape in this paper eans complex geometric features, for example, the shape of an ccentric bearing, end relief, and so on. It is a very complemenary work to Fantino et al. 2 in the way that the effects of other arameters are analyzed. It is shown that some bearing geometry efects have a significant influence on minimum film thickness. antino and Fr\u00eane 8 studied the influence of the engine type petrol and diesel on the same result, but no conclusion could be ade about the impact of other parameters load and speed Howver, as concerns the speed influence, an EHD parametric result an be found in Ref. 9 , which shows a decrease of the minimum lm thickness in the range 100\u2013700 rpm. A more complete paraetric study, involving the load, engine speed, and bearing stiffess, is reported in Ref. 10 Later, Bonneau and Guines 11 developed a FEM based algoithm to calculate the EHD behavior of connecting rod big-end earings, taking into account the body force deformation and caviation effects. Boedo and Booker 12,13 investigated, under isohermal conditions, the coupled effects of body force deformation nd mass-conserving cavitation effects on the EHD behavior of onnecting rod big-end bearings.\nFantino et al. 4,14 focused on the effect of the viscosity on the inimum film thickness for a connecting rod big-end bearing and shijima et al. 15 developed an EHD model showing that oil iezoviscosity has an influence on bearing performance.\nThe effects of bearing length and housing stiffness on the conecting rod big-end bearing have been reported by Okamoto et al. 16 . The results proved that the decrease of the bearing length has significant incidence on the minimum film thickness and maxium pressure due to the load capacity reduction. The thermoelastic deformations play an important role for the onnecting rod bearing transient lubrication because load direcion, as well as load modulus, strongly varies in time. Experimenal works by Goodwin and Holmes 17 , Conway-Jones and Goon 18 , Hashizume et al. 19 , and Suzuki et al. 20 show that hermal effects in engine bearings are actually essential.\nFrom this abundant and interesting literature, the following rearks can be made.\n1 Several parameters have been reported as \u201cimportant\u201d for the connecting rod big-end bearing behavior: the lubricant viscosity, the bearing geometry, and the operating conditions. 2 No parameter has been really classified as noninfluent or of minor influence. 3 The effects of the majority of these parameters have been\nassessed independently.\n11101-2 / Vol. 131, JANUARY 2009\nom: http://tribology.asmedigitalcollection.asme.org/ on 01/28/2016 Terms\nHence, the main question we are interested in has no answer for the moment: What is the effect of a parameter compared with another? Are some of theses parameters nonsignificant and therefore of little interest? To try to answer to these questions, the \u201ctools\u201d used in the present paper will be presented.\n3 Numerical Code Presentation The numerical code, used in the present work, is particularly dedicated to lubricant film characteristics determination and especially big-end connecting rod bearings lubrication. Given the following input parameter families:\n1 connecting rod meshed computer-aided design CAD , Figs. 1 and 2, with its physical properties 2 bearing shell geometry including defects and physical properties 3 engine speed and applied load variations 4 lubricant properties 5 lubricant supply properties location, pressure, and\ntemperature\nthe nonexhaustive list of computed physical values is as follows:\na film pressure and thickness fields b power loss power dissipated by the lubricant shearing c operating temperature d flow rate\n4 Elastohydrodynamic Governing Equations Using the usual assumptions of lubrication theory and considering a laminar flow with neglected inertia effects in the film, the Reynolds equation, in the incompressible case, may be written as follows:\nx G\np x + y G p y = U F x + h t 1\nwhere\nF = J1\nJ0\nG = J2 \u2212 J1\n2\nJ0\nand\nTransactions of the ASME\nof Use: http://www.asme.org/about-asme/terms-of-use",
+            "n i\nw s a d e\nT f\nW\nw a h f n p b\nw a p\nw p\nfl i t t\nJ\nDownloaded Fr\nJi = 0\nh i\nd\nThis equation can only be solved for the full film zones. For the on-active cavitation film zones, a second equation must be ntroduced:\nU h\nx + 2\nh t = 0 2\nhere is the density of the lubricant-gas mixture. In order to olve Eqs. 1 and 2 simultaneously, a universal variable D and n effective film thickness variable r are defined. If 0 is the oil ensity, the latter is given by r= h / 0. A generalized Reynolds quation can thus be written:\n2B x G\nD x + 2B y G D y\n= 2U F\nx + 2\nh t + 1 \u2212 B D x + 2 D t 3\nhe universal variable D and the cavitation index B are defined as ollows.\n1 In full active zone,\nD = p \u2212 pcav, D 0\nB = 1\n2 In cavitated non-active zone,\nD = r \u2212 h, D 0\nB = 0\nithout misalignment, the film thickness is given by\nh ,y,t = h0 ,t + he ,y,t + ht ,y,t + hd ,y,t 4\nhere =x /R is the angular coordinate for a housing of R radius nd h0 , t is the nominal film thickness; for a circular bearing 0 , t =c 1\u2212 x t cos \u2212 y t sin , he ,y , t is the elastic deormation of the bearing housing and shaft due to the hydrodyamic pressure, ht ,y , t is the deformation due to thermal exansion of the bearing housing and shaft, and hd ,y , t defines a arrel and/or a lemonlike shape, explained on Fig. 3.\nThe lubricant viscosity is assumed to vary with the temperature\nT = 0e\u2212 T\u2212T0 + a 5\nhere a is an asymptotic viscosity, 0 is the oil viscosity at T0, nd is the thermoviscosity coefficient. It is also supposed to be ressure dependent according to the Barus equation, and then\nP,T = T e P\u2212P0 6\nhere is the piezoviscosity coefficient and P0 is the ambient ressure.\nTo determine the mean temperature T, the power resulting from uid shearing Pf is considered to be 80% evacuated by the fluid\ntself value commonly admitted for engine bearings and 20% hrough the solids. The fluid temperature rise T, above the inlet\nemperature, is then obtained with the following relationship:\nournal of Tribology\nom: http://tribology.asmedigitalcollection.asme.org/ on 01/28/2016 Terms\nCpQ T = 80%Pf 7\nwhere is the oil density, Cp is the oil specific heat capacity, and Q is the flow rate. Data for typical engine oils 21 up to pressures of 200 MPa indicate that reduces with oil film pressure P MPa and with temperature T \u00b0C , in a manner which can be represented by\n= 0.139P\u22120.13T\u22120.36 8 This relationship will be useful to determine the variation range of taking into account the thermal and pressure effects.\nThe balance of the applied loads with the hydrodynamic pressures leads to\nS\np cos dS \u2212 Fx = Mx\u0308\nS\np sin dS \u2212 Fy = My\u0308 9\nwhere Fx and Fy are the applied loads acting on the bearing, M is the connecting rod mass, and x ,y are the connecting rod center of mass coordinates.\nThe boundary conditions used to solve the modified Reynolds equation are based on the active/non-active film zone separation and are detailed by Hajjam and Bonneau 22 .\n5 Elastic Deformations The elastic component in the film thickness equation is due to the hydrodynamic pressure acting on the housing and the shaft. Because connecting rod bearings are highly compliant, the radial displacement at a given point depends on the whole pressure field. Using finite element discretization, compliance matrices Ce\nh and Ce\ns are then calculated: A radial unit force is successively applied at each surface node of the housing and shaft, respectively. Radial distortion under these unit loads gives elementary solutions, compiled in the compliance matrices. For the housing compliance, it is assumed that the connecting rod is fixed at some distance from the bearing. The mean bending displacements are then subtracted from the housing distortion, in order to eliminate the fixation effect. The elastic deformation, at a given node i, is as follows:\nhe i =\nk=1\nnns\nCe s i,k fk +\nk=1\nnnh\nCe h i,k fk 10\nwhere the applied force fk on node k is obtained by integration of the pressure field.\n6 Thermal Deformations The modification of the film thickness due to the thermal deformation has two origins. The first one is the shaft thermal expansion and the second one is the housing growth. The thermal compliance matrices are constructed in the same way as the elastic ones thus, for a given node i, all of the node k contributions are summed but Tk kept constant fluid mean temperature :\nht i =\nk=1\nnns\nCt s i,k Tk \u2212 Tref \u2212\nk=1\nnnh\nCt h i,k Tk \u2212 Tref 11\nThis procedure was first introduced by Khonsari and Wang 23 and then widely used.\n7 FEM Formulation for the EHD Problem The FEM formulation for the EHD problem is fully detailed in Ref. 22 . However, for a better understanding of the present paper, a short reminder is given:\na Two different problems must be formulated.\nJANUARY 2009, Vol. 131 / 011101-3\nof Use: http://www.asme.org/about-asme/terms-of-use"
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0000373_50006-8-Figure6.24-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000373_50006-8-Figure6.24-1.png",
+        "caption": "Fig. 6.24. Effect of Turning Tool Nose Radius on Workpiece 4~",
+        "texts": [
+            " The side and end relief angles, which provide clearance between the tool flank and the workpiece, are usually a compromise: too small of an angle will cause rubbing, while too large of an angle will provide insufficient cutting edge strength. The side cutting angle affects the load on the cutting edge and provides thickness and directional control to the chips. The nose radius provides strength to the tool nose and helps to dissipate heat generated by the cut. A scallop produced by a tool with a nose radius (Fig. 6.24) gives a better surface finish, shallower scratches, and a stronger workpiece, with less tendency to crack at sharp comers than the notch effect produced by a sharp tool. Water soluble oils in mixtures (1 part oil with 20-40 parts water) are frequently used in turning. Flood cooling is used to minimize excessive heat build-up. If sulfurized or chlorinated oil is used as a cutting fluid, the workpieces must be thoroughly cleaned before heat treatment or high temperature service. Serious damage to workpieces during heating cycles can result if any residue remains"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002745_20110828-6-it-1002.02694-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002745_20110828-6-it-1002.02694-Figure1-1.png",
+        "caption": "Fig. 1. Proposed UAV Model from Different Angle",
+        "texts": [
+            " However, the analysis of controllabilily indicates that other meaningful motions can be realized. Taking the results of model analysis into account, a nonlinear controller is designed in section 4 with the brief introduction of MIMO Output Zeroing Control. Finally, numerical simulation verifies the performance of position control along with the upward posture. In this section, the two nonlinear state equations, Euler angle based model and quaternion based model, are derived for the proposed Trirotor UAV model shown in Fig.1 with the consideration of results by T. Cheviron and Plestan (2009). One of the main features of this model is the three rotors which is installed on three skew axes. One note here is that this model does not own other inputs such as a servo motor to control tilting angle as modeled. Since the main intention of this paper is the analysis and control design 978-3-902661-93-7/11/$20.00 \u00a9 2011 IFAC 10391 10.3182/20110828-6-IT-1002.02694 for an underactuated UAV model, this model can be a good starting example"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003741_1754337111425629-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003741_1754337111425629-Figure3-1.png",
+        "caption": "Fig. 3 Pendulum arrangement for Moment of Inertia measurements",
+        "texts": [
+            " Under this scaling condition the modal mass can be viewed as the amount of mass that is participating in each mode of vibration. Since the structure was freely suspended during the test, then the low frequency residuals can be a direct measure of rigid body mass properties of the structure. The scaled mode shapes were used to calculate the energy absorbed at each excitation point on the blade maintaining constant the position of the accelerometer [3]. The moment of inertia about an axis through the handle of the bat was determined by turning the bat into a physical pendulum (Fig. 3). The period of a physical pendulum is given by the formula: T = 2 p ffiffiffiffiffiffiffiffiffiffi I mgd s \u00f01\u00de where I is the moment of inertia of the bat, m is its mass, and d is the distance between the pivot point and the centre of mass (CoM). A 0.006 m hole was drilled through the handle of the bat and a rod was inserted through the hole to form an axis of rotation. A frame was manufactured out of mild steel to support the rod so as to allow the bat to be held in a vertical position in order to swing freely"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003727_j.engfailanal.2011.06.004-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003727_j.engfailanal.2011.06.004-Figure1-1.png",
+        "caption": "Fig. 1. Schematic representation of the cooling fan assembly, U-bolt and fracture position.",
+        "texts": [
+            " Many factors such as variation of the centrifugal stress, shut up/down of the cooling system and mass imbalance make the fatigue to be the most probable cause of failure in cooling fan systems. It is reported that about 90% of all mechanical failures is caused by fatigue [2]. Fatigue is one of the most dangerous mechanical failures because it occurs under loads that are lower than the static strength of the material [3]. The case studies on the fatigue fracture can be found in many sources such as Refs. [4\u20137]. In the present study, fatigue fracture of holding Ubolt of a cooling fan is analyzed using fractography examinations and finite element modeling. Fig. 1 shows a schematic drawing of the fan blade. A holding U-bolt and fracture position are also schematically presented in Fig. 1. The fracture position is near the nuts and below the hub plate. The chemical composition of the U-bolt material is presented in Table 1. This indicates that the bolts are made by 304-type stainless steel. Typical values of ultimate tensile stress and yield stress for this type of stainless steel are given as 505 and 205 MPa, respectively [8]. The microstructure of the U-bolt was examined by cutting a sample from the fractured bolt, mechanical grinding and polishing to a mirror-like surface using alumina powder solution",
+            " In the stress-based approach, the effect of mean stresses in multiaxial fatigue is accounted for by using mean stress plots such as Goodman diagram. The equivalent mean stress, Sme, for multiaxial fatigue can be defined as the sum of principal mean stresses as Sme \u00bc Sm1 \u00fe Sm2 \u00fe Sm3 \u00f08\u00de where Sm1, Sm2 and Sm3 are the principal mean stresses. By replacing Sa and Sm in Eq. (5) by Sae and Sme, respectively, the fatigue stress is estimated for multiaxial fatigue. Principal alternating and mean nominal stresses are calculated for all nodes using FE software. Referring to Fig. 1, two end parts of U-bolts are threaded to permit the nuts to be screwed up to the hub plate bottom. Therefore, fatigue stress concentration due to threaded part raises the effective nominal stress amplitude, Sae. Fatigue stress concentration factor, Kf, is obtained from notch sensitivity factor, q, and stress concentration factor, Kt, by K f \u00bc 1\u00fe q\u00f0Kt 1\u00de \u00f09\u00de In order to choose the values of Kt and q, the values of the models used in Lehnhoff and Bunyard [15] work that has the closest diameter to the present work are chosen here as 4"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003779_s11071-012-0647-0-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003779_s11071-012-0647-0-Figure5-1.png",
+        "caption": "Fig. 5 Inverted pendulum system; an inverted pendulum is pinned to a moving cart",
+        "texts": [
+            " When the pendulum is raised from the rest position to its upright one, the system is strongly nonlinear with the pendulum angle. The input to the system is the force applied to the cart, which is free to move. In this section, we apply the fuzzy-Pad\u00e9 controller to the single inverted pendulum system, and then the results are compared to a fuzzy controller with a similar rule base. 3.1 System dynamics and the fuzzy controller The single inverted pendulum system is composed of two bodies: a cart and a pendulum (Fig. 5). The pendulum is attached to the cart by a revolute joint and the goal of the controller is to stabilize the pendulum in its unstable upright position without taking the cart position into consideration. The dynamic model of this second-order system in the state space form can be written as [39] [ x\u03071 x\u03072 ] = \u23a1 \u23a3 x2 g sinx1\u2212kmlx2 2 sin(2x1)/2+ku cosx1 4l/3\u2212kml cos2 x1 \u23a4 \u23a6 , (11) where k = 1/(M + m). In the above equation, x1 = \u03b8 and x2 = \u03b8\u0307 are the rotation angle and angular velocity of the pendulum, respectively"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002793_1.4001214-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002793_1.4001214-Figure6-1.png",
+        "caption": "Fig. 6 Simulation of the gear meshing with assembly errors",
+        "texts": [
+            " 5 Simulation of the Condition of Meshing and Contact The TCA technique has also been applied to simulate the condition of meshing, whose main goal is to determine the shift of the bearing contact and the transmission errors caused by the misalignment of the gear drive. The first step in such application is to represent the equations for two mating tooth surfaces in the same coordinate system, whose coordinates and unit normal should be the same at the point of contact. Then the kinematic characteristics of this type of gearing can be analyzed by combining TCA with the mathematical model. After setting Sp xp ,yp ,zp , Sg xg ,yg ,zg , Sf xf ,yf ,zf , and Sh xh ,yh ,zh , as shown in Fig. 6, the conditions of meshing can be simulated by changing the settings and the orientations of the coordinate system St with respect to Sf. The coordinate systems St and Sh are used to simulate the errors in the alignment, such as the axial displacement E, the change in the center distance L, and the axis misalignments and . Coordinate transformations can then be used to produce the following equations of the gear tooth surface Re g and the pinion tooth surface Re p in the fixed coordinate system Sf R f p = M fh \u00b7 Mht \u00b7 Mte \u00b7 Re p 13 and R f g = M fe \u00b7 Re g. 14 As shown in Fig. 6, when two gears mesh with each other, the center distance is L =L+ L, and p and g represent the angles of rotation of the pinion and gear, respectively. Likewise, the tan- utter heads and an imaginary genhelical gear o c vex gency of the two contacting tooth surfaces means that they have a Transactions of the ASME 6 Terms of Use: http://www.asme.org/about-asme/terms-of-use c p s a S E t T s J Downloaded Fr ommon position vector and the same surface unit normal at the oint of contact. The contacting surfaces and their unit normal in the coordinate ystem Sf can be represented by the following equations: R f p jkp, Hkp, Gp, p , L, E, , = R f g jkg, Hkg, Gg, g 15 n f p jkp, Hkp, Gp, p , , = n f g jkg, Hkg, Gg, g 16 f f p jkp, Hkp, Gp = 0 17 nd f f g jkg, Hkg, Gg = 0 18 ince Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001891_1.3159377-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001891_1.3159377-Figure5-1.png",
+        "caption": "Fig. 5 Reference coordinate system for rotating tests with shaker loads",
+        "texts": [
+            " This behavior is attributed to the repositioning of the bristles and the stiffening effect due to the pressure differential across the seal i.e., blowdown effect . Presently, in the rotordynamic measurements with a centered seal, the gas supply pressure Ps is manually adjusted 169 kPa and 238 kPa , and the motor is turned on to bring the test rotor to a constant speed 600 rpm and 1200 rpm . As the shaft spins, the electromagnetic shaker excites the test seal with a periodic load, amplitude of 22 N,2 and single frequency ranging from 20 Hz to 80 Hz 3 Hz increments . Figure 5 shows the reference coordinate 1Pressure ratio Pr= Ps / Pd=absolute supply pressure/absolute discharge pressure. 2Smaller load magnitudes lead to stick-slip nonlinear phenomenon with erratic seal behavior, while larger magnitude loads produce too large rotor displacements that endanger the seal life. operties of test hybrid brush seal SI Unit U.S. unit 167.1 mm 6.580 in. 166.4 mm 6.550 in. 183.1 mm 7.210 in. 8.53 mm 0.336 in. 0.381 mm 0.015 in. 20 7.23 mm 0.331 in. 45 deg - 22.48 105 bars 32.6 106 psi 850 bristles/cm 2300 bristles/in"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001891_1.3159377-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001891_1.3159377-Figure4-1.png",
+        "caption": "Fig. 4 Detail view of disk/shaft assembly",
+        "texts": [
+            " References 17,19 report the measurements of seal leakage and power loss and the identification of static structural stiffness, respectively. 2 Description of Test Rig and HBS Figure 3 depicts the HBS rotordynamic test rig and its instrumentation. A slender steel shaft is affixed to the base of a cylindrical steel vessel via two tapered roller bearings. The free end of the shaft holds a steel disk where the test seal is located. Two eddy current sensors, 90 deg apart, are secured on the front plate of the vessel and face the outer diameter of the steel disk, as shown in Fig. 4. The sensors record the disk displacements along two orthogonal directions in the vertical plane. A slender rod stinger with load cell connects an electromagnetic shaker to the free end of the shaft. Two soft springs located at the drive end of the shaft, in the vertical and horizontal directions, allow the centering of the rotor free end with respect to the test seal. The springs connect to the shaft through a ball bearing. At the shaft\u2019s free end, a small dc motor drives the overhang shaft and disk Transactions of the ASME 6 Terms of Use: http://www"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001174_cae.20245-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001174_cae.20245-Figure1-1.png",
+        "caption": "Figure 1 Sketch of the 14 degree of freedom vehicle model.",
+        "texts": [
+            " Furthermore, longitudinal and transversal bars connect both masses in order to achieve a better control of roll and pitch movements. Finally, according to Ref. [18], the consideration of the anti-roll bar provides significant improvement to the model results and the computational cost of this additional operation is quite low. This bar transmits forces to the sprung and unsprung masses. These forces are proportional to the roll angle and have the opposite direction, so the final effect is an additional rolling stiffness. Figure 1 shows a schematic view of the vehicle model. Dynamic Equations of Sprung Mass Movements. Model equations have been written considering the Cartesian coordinates system fixed to the sprung mass that Figure 2 shows. Linear and angular movement equations are found. It should be noted that a non-inertial coordinates system is used, so inertial forces appear. External forces include aerodynamic, tire and suspension forces. Linear movements: FX \u00bc m du dt \u00fe qw rv \u00f01\u00de FX \u00bc m dv dt \u00fe ru pw \u00f02\u00de FX \u00bc m dw dt \u00fe pv qu \u00f03\u00de Angular movements MX \u00bc IX dp dt \u00f0IY IZ\u00deqr \u00fe IXY rp dq dt IYZ\u00f0q2 r2\u00de IZX pq \u00fe dr dt \u00f04\u00de MY \u00bc IY dq dt \u00f0IZ IX\u00derp \u00fe IYZ pq dr dt IZX\u00f0r2 p2\u00de IXY qr \u00fe dp dt \u00f05\u00de MZ \u00bc IZ dr dt \u00f0IX IY\u00depq \u00fe IZX qr dp dt IXY\u00f0p2 q2\u00de IYZ rp \u00fe dq dt \u00f06\u00de where FX, FY, and FZ are the external forces, MX, MY, and MZ are external momentums, u, v, and w are linear velocities, p, q, and r are angular velocities around the three axes of the reference system, m the vehicle mass, IX, IY, and IZ are moments of inertia of the vehicle around the three axes of the reference system, and IZX, IXY, and IYZ are moments of inertia of the vehicle"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003658_09507116.2011.600013-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003658_09507116.2011.600013-Figure1-1.png",
+        "caption": "Figure 1. Schematic set-up of laser beam, welding wire, and filler in FLA welding.",
+        "texts": [
+            " In this article, we discuss the gap tolerance of butt welds of thin plate or thick plate and the problem of enlarging this, describe a study of the basic phenomena of bead formation and filler fusion during FLA welding and the effects of enlarging the gap tolerance when this method is used for butt welding. There is also a discussion of poor bead appearance that is likely to occur in butt welds of thin or thick plates. ISSN 0950-7116 print/ISSN 1754-2138 online q 2012 Taylor & Francis http://dx.doi.org/10.1080/09507116.2011.600013 http://www.tandfonline.com Welding International Vol. 27, No. 2, February 2013, 98\u2013108 Selected from Journal of Light Metal Welding and Construction 48(11) 424\u2013434 The set-up of the laser beam, welding wire, and filler is shown in Figure 1. The laser beam and welding wire are arranged at a suitable distance so that the transfer of droplets from the wire tip to the molten pool does not interfere with the laser. The gradients of these aL and aA should be such as to facilitate the set-up and care must be taken to ensure that light reflected from the molten pool does not return to the focusing system. The filler is arranged at a fixed angle aF and in such a way that it is directly melted by the laser beam. Welding is carried out with the welding wire constantly sloping forward to ensure a good bead appearance22,23",
+            " At 120 A and above, a good weld was achieved by both pulsed MIG arc welding and FLA welding. The maximum deposition rate of FLA welding at which a good bead appearance was obtained is shown in Figure 10 accompanied by that for pulsed MIG arc welding. With pulsed MIG arc welding, the deposition rate increase was almost completely in proportion with the weld current. With FLA welding, it was possible to obtain a weld speed two- to threefold that achieved by pulsed MIG arc welding. The above results are from using 2-mm-thick A5052 aluminium alloy sheet. Since, as in the set-up shown in Figure 1, the filler is directly irradiated by the laser beam, it is thought that the deposition rate is not affected by plate thickness. It is possible, however, that filler fusion is affected by the arc and this is discussed in Section 4. As noted in Section 3.2.1, when FLA welding is carried out, it is possible to adjust the deposition rate over a wide range without increasing the welding current. In the set-up shown in Figure 1, in addition to the increase in deposition rate, the filler also contributes to the prevention of the laser beam escaping from the gap as it is first applied during welding of a butt joint with a gap. Accordingly, examination was made of the enlargement of gap tolerance in butt joints of 2-mm-thick A5052 aluminium alloy sheet and 8-mm-thick A5052 and A6061 aluminium alloy sheets. The results of the examination of butt joints of 2-mmthick A5052 aluminium alloy sheet are described as follows. The bead appearance and penetration cross section, with the gap and welding speed varied, of butt joints made with a 4 kW fibre laser at a welding current of 115 A are shown in Figure 11(a) and (b)",
+            " It is possible to raise the deposition rate two- to threefold that of hybrid welding without increasing the welding current and the dimple in the molten pool surface, which occurs with hybrid welding, is filled by the deposited metal from the filler. Even with FLA welding, however, a concavity does appear in the molten pool surface when the gap reaches 3 mm, as shown in Figure 12(b). This is thought to be because the gap is too wide for the filler supplied. In the set-up used in this study, and as shown in Figure 1, the filler is positioned so as to be directly melted by the laser and at a specific distance from the pulse MIG arc. It is unclear from an inspection of this set-up diagram alone whether the filler is melted by the arc. It is clear, however, from the relationship between deposition rate and welding current shown in Figure 10, the maximum filler deposition rate increases with an increase in the welding current up to a current of 150 A even at the same laser power. This shows that the arc contributes to the filler fusion in some form"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000365_cacsd-cca-isic.2006.4776641-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000365_cacsd-cca-isic.2006.4776641-Figure1-1.png",
+        "caption": "Fig. 1. Workspace W with car body A and obstacles Bi",
+        "texts": [
+            " Moreover, some basic consequences resulting from the special structure and fundamental properties of the car model are stated. In Section III the principles of the proposed two-step approach are explained, before the efficient implementation is addressed in Section IV. Finally, a simulation result is presented in Section V followed by some concluding remarks in Section VI. First, the problem stated during the introduction is specified more precisely applying basic tools and notions from robotics literature (see e.g. [1]). As shown in Figure 1 the body of a car A is approximated as a compact set of points within a rectangular region moving in a two-dimensional Euclidean space, which is referred to as workspace W . Analogously, obstacles in the neighborhood of the parking space are represented as the compact sets of points Bi, i = 1, 2, . . . , m. Assuming small velocities the dynamics of a car having axle base L can be described by a single-track model satisfying the state equations \u23a1 \u23a2\u23a2\u23a3 x\u0307 y\u0307 \u03b8\u0307 \u03c6\u0307 \u23a4 \u23a5\u23a5\u23a6 = \u23a1 \u23a2\u23a2\u23a3 cos \u03b8 sin \u03b8 1 L tan \u03c6 0 \u23a4 \u23a5\u23a5\u23a6 v + \u23a1 \u23a2\u23a2\u23a3 0 0 0 1 \u23a4 \u23a5\u23a5\u23a6 \u03c9 (1) 0-7803-9796-7/06/$20.00 \u00a92006 IEEE 163 with the state variables as defined in Figure 1. The inputs v and \u03c9 are the car velocity at the point [x, y]T and the angular steering velocity, respectively. Furthermore, the steering angle and angular velocity are constrained according to |\u03c6| \u2264 \u03c6max < \u03c0 2 and |\u03c9| \u2264 \u03c9max. The system description (1) can be further simplified by introducing the new state \u03ba = 1 L tan \u03c6 and the new input \u03c3 = 1 L cos2 \u03c6\u03c9 which yields \u23a1 \u23a2\u23a2\u23a3 x\u0307 y\u0307 \u03b8\u0307 \u03ba\u0307 \u23a4 \u23a5\u23a5\u23a6 = \u23a1 \u23a2\u23a2\u23a3 cos \u03b8 sin \u03b8 \u03ba 0 \u23a4 \u23a5\u23a5\u23a6 v + \u23a1 \u23a2\u23a2\u23a3 0 0 0 1 \u23a4 \u23a5\u23a5\u23a6\u03c3 (2) with the configuration1 vector q = [x, y, \u03b8, \u03ba]T and the restrictions |\u03ba| \u2264 \u03bamax = 1 L tan \u03c6max (3) |\u03c3| \u2264 \u03c3max = 1 L \u03c9max \u2264 1 L cos2 \u03c6 \u03c9max (4) Note that in (4) the resulting state-dependent constraint of the new input \u03c3 is estimated by a lower state-independent bound, which is reasonable as long as cos2 \u03c6 \u2248 1, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000928_2008-01-1019-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000928_2008-01-1019-Figure1-1.png",
+        "caption": "Figure 1: Functional structure of a dual mass flywheel",
+        "texts": [
+            " In-cylinder pressure sensors are, due to their low cross-sensitivity, expensive and predominantly used in luxury vehicles with 6 or more cylinders. Acceleration transducers have also been used to detect cylinder torque variations [3]. In order to provide an acceptable solution for low and midcost vehicles, cylinder balancing systems have to fulfil demanding robustness requirements at very low cost. As previously noted the crankshaft speed sensor method is very sensitive with respect to driveline torque reactions. This combined with the now widespread utilisation of the Dual Mass Flywheel (DMF), shown in Figure 1, has provoked a search for more robust and low cost methods of detecting and correcting cylinder torque imbalances. Basically, the DMF includes two large rotating flywheel inertias connected by long travel arc-springs. More detailed information about the DMF can be found in [4]. One of the principal advantages of the DMF is the almost complete elimination of secondary flywheel side and thus gearbox input-shaft speed oscillations. Annoying gear rattle as well as other noises and vibrations are eliminated, allowing the engine to be driven at lower speeds, improving real world fuel economy"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000546_s12239-008-0072-z-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000546_s12239-008-0072-z-Figure6-1.png",
+        "caption": "Figure 6. Generation of an S-path using two arcs of maximum radius.",
+        "texts": [
+            " The radius of the arc, R2, is calculated as follows; (4) where, l1=Py The coordinate position of point B (denoted as B(bx, by)) at which a straight travel path starts, is calculated as R1= a sin\u03b8 ---------- tan 180o \u03b8\u2013 2 -------------------\u239d \u23a0 \u239b \u239e \u03c8 = 180 \u03b8\u2013( ) 2 ---------------------- R2= l1( )tan 180o \u03b8\u2013 2 -------------------\u239d \u23a0 \u239b \u239e follows: PC=PB=l1=Py Bx=PD=PB \u00d7 cos(90\u2212\u03b8)=l1 cos(90\u2212\u03b8) By=l1+BD=l1+l1 sin(90\u2212\u03b8) (5) Case 2-1 (Py<0 and \u03b8 >0) If Py<0, a path of one arc is not sufficient to get to the target position and meet the target direction. In this case, two arcs of different direction should be employed as revealed in Figure 6. Figure 6 shows a typical S-path. In a true sense, this S path is actually an inverted version of an S-path. After some extensive algebraic manipulation, the equation for the derivation of R3 is given in Equation (6). To maintain the clarity of the paper, the detailed steps of derivation have been omitted and so have the steps to calculate the coordinate position of A, A(Ax, Ay), at which a different arc path starts. (6) Case 2-2 (Py<0 and \u03b8=0) A special case arises when the target direction is parallel to the current direction, i"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002487_j.precisioneng.2011.02.001-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002487_j.precisioneng.2011.02.001-Figure4-1.png",
+        "caption": "Fig. 4. Center of curvature of an aspherical surface at radial distance y1.",
+        "texts": [
+            " The optical axis of an aspherical element is formed by joining ll the centers of curvature together. The center of curvature at a ertain radial distance is necessary for the following assembly simlation. The radial distance is usually the supporting diameter of he mounting shoulder or the contacting circular diameter of the pacer. As the surface is rotationally symmetrical about the optical xis, the center of curvature at an arbitrary s value can be deterined by selecting a cross-section, for example x = 0, as illustrated n Fig. 4. The radial distance from the z-axis is s2 = y2. Therefore, the lope of the curve at (y1, z1) is: an = 2 \u00d7 y f (y1 + Dy) \u2212 f (y1 \u2212 Dy) (11) here y is an infinitesimal increment of y1. As a result, the oordinates of the center of curvature are (0, 0, z1 + y1 tan ). As ach aspherical surface has its own optical axis, the surface tilt nd decenter are defined separately in the specification list. The etailed descriptions can be found in the following example in Secion 4. A biconvex aspherical lens is shown in Fig"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001219_s12541-009-0102-4-Figure6-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001219_s12541-009-0102-4-Figure6-1.png",
+        "caption": "Fig. 6 Load application points and displacement measurement points for estimating the normal, lateral, and axial compliances of the X-axis feed system",
+        "texts": [],
+        "surrounding_texts": [
+            "A virtual prototype of the ultra-precision machine for machining large-surface micro-features, which was constructed based on ANSYS software to estimate its compliances, is presented in Fig. 10. The virtual prototype was composed of 134,729 nodes, 534,911 solid elements, and 408 matrix elements. The matrix elements were introduced in order to represent the normal and lateral stiffnesses of the hydrostatic guideways, the radial and thrust stiffnesses of the hydrostatic bearings, and the axial stiffness of the linear motors. As the boundary condition for the structural analysis and measurement, the movement of the ultra-precision machine was restricted in the vertical direction at the four supporting points of the bed, as shown by the red arrows in Fig. 10. The bed, column, cross beam, and feed tables were made of cast iron (GC300), the C-axis shaft and bracket were made of steel (SCM440, SS400), and the mover/rotor and stator of motors were made of Fe-Si. Table 3 shows the material properties used for the virtual prototype."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001681_12.851654-Figure8-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001681_12.851654-Figure8-1.png",
+        "caption": "Figure 8. Multibody system of human body",
+        "texts": [
+            "org/terms In the final process, after detection and tracking procedures of markers have been done, by using calibration data, the marker\u2019s coordinates in the real world can be obtained. Finally, using this data, some gait parameters, such as joint angle, step length and gait velocity can be calculated. To conduct dynamic analysis of human gait, in this work a multibody system is constructed to model the human body. The body is modeled by a system of five rigid bars connected by joints12 as shown in Fig. 8. The five bars represent two legs, two thighs, and a body that consist of head, arms and trunk13. The bars are having mass concentrated at its centroid denoted by B1, B2, B3, B4 and B5, where the mass of each bars is following the anthropometric data provided by Winter14. The ankles are represented by points A and F, while B and E are the knees. Point C is the hip and D is the head. With this model, the mathematical formulation for dynamics analysis can be constructed 12-13, 15-16. A program is written in Matlab R2007b to plot and calculate the kinematics and dynamics data of human gait"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000042_50001-0-Figure1.20-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000042_50001-0-Figure1.20-1.png",
+        "caption": "Fig. 1.20. Construction of stability boundary (upper diagram, from Fig. 1.17). On the isolated",
+        "texts": [
+            "84) may be violated when we deal with tyre characteristics showing a peak in side force and a downwards sloping further part of the characteristic. The second condition corresponds to condition (1.65) for the linear model. Accordingly, instability is expected to occur beyond the point where the steer angle reaches a maximum while the speeA is kept constant. This, obviously, can only occur in the oversteer range of operation. In the handling diagram the stability boundary can be assessed by finding the tangent to the handling curve that runs parallel to the speeA line considered. In the upper diagram of Fig. 1.20 the stability boundary, that holds for the right part of the diagram (ay vs l/R), has been drawn for the system of Fig. 1.17 that changes from initial understeer to oversteer. In the middle diagram a number of shifted V-lines, each for a different steer angle fi, has been indicated. In each case the points of intersection represent possible steady-state solutions. The highest point represents an unstable solution as the corresponding point on the speeA line lies in the unstable area. When the steer angle is increased the two points of intersections move towards each other",
+            "6): Fy = D sin[C arctan{Ba - E ( B a - arctan(Ba)) } ] We define: the peak side force D= luF z and the cornering stiffness CFa = BCD=cF~F z so that B = CFa/(CI~). For the six tyre/axle configurations the parameter values have been given in the table below. axle case p front a, b 0.8 c 0.78 a 0.9 rear b 0.9 c 0.65 C Fa 11 11 C E 1.2 -2 1.3 -2 1.2 -2 1.2 -2 1.5 -1 Determine for each of the three combinations (two dry, one wet): 4. The stability boundary (associated with these oversteer ranges) in the (ay /g v e r s u s l/R) diagram (= right-hand side of the handling diagram) (cf. Fig. 1.20). 5. Indicate in the diagram (or in a separate graph): a. the course of the steer angle 0 required to negotiate a curve with radius R = 60m as a function of the speed V. If applicable, indicate the stability boundary, that is the critical speed Vcrit , belonging to this radius. b. the course of steer angle 0 as a function of relative path curvature l/R at a fixed speed V = 72 km/h and if applicable assess the critical radius Rcrit\" For the vehicle systems considered so far a unique handling curve appears to suffice to describe the steady-state turning behaviour",
+            "22 appears to be open on two sides which means that initial conditions, in a certain range of (r/v) values, do not require to be limited in order to reach the stable point. Obviously, disturbance impulses acting in front of the centre of gravity may give rise to such combinations of initial conditions. In Figs. 1.23 and 1.24 the influence of an increase in steer angle ~ on the stability margin (distance between stable point and separatrix) has been shown TYRE CHARACTERISTICS AND VEHICLE HANDLING AND STABILITY 49 for the two vehicles considered in Fig. 1.20. The system of Fig. 1.23 is clearly much more sensitive. An increase in 6 (but also an increase in spee~ V) reduces the stability margin until it is totally vanished as soon as the two singular points merge (also the corresponding points I and II on the handling curve of Fig. 1.17) and the domain breaks open. As a result, all trajectories starting above the lower separatrix tend to leave the area. This can only be stopped by either quickly reducing the steer angle or enlarging 6 to around 0.2rad or more"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001685_s11668-009-9233-2-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001685_s11668-009-9233-2-Figure1-1.png",
+        "caption": "Fig. 1 Schematic illustration of the lorry wheel hub system",
+        "texts": [
+            " In addition two-dimensional (2D) finite element analysis was carried out to determine the stress distribution on inner ring surface. Tapered roller bearings are generally separable; that is, the cone, consisting of the inner ring with roller and cage assembly, can be mounted separately from the cup (outer ring). The surface finish of the tracks and rolling elements is critical to the running performance and noise and vibration characteristics of these bearings. An example for a tapered roller bearing used in a wheel hub is given in Fig. 1. Tapered roller bearings particularly are suitable for the accommodation of combined (radial and axial) loads. The axial load-carrying capacity of the bearings is largely determined by the contact angle, the larger contact angle, the higher the axial load-carrying capacity [3]. Y. Kayal\u0131 Department of Metal Education, University of Afyon Kocatepe, Afyon, Turkey I. Ucun K. Aslantas\u0327 (&) Department of Mechanical Education, University of Afyon Kocatepe, Afyon, Turkey e-mail: aslantas@aku.edu.tr Bearing Characteristics The tapered rolling bearing considered in this work was designed to withstand axial and radial loads"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003059_s12206-011-1027-2-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003059_s12206-011-1027-2-Figure2-1.png",
+        "caption": "Fig. 2. Collision between a point mass and a rigid rod.",
+        "texts": [
+            " 3 2 sbf ay by cy= + + (10) where fsb is transverse force of the string bed in the center, y is lateral deflection of the string bed in the center, and a, b and c are system parameters determined by the optimization technique [5]. It should be noted that the slope of the function, fsb, represents the transverse stiffness of the string bed. In this study, the impact point of a ball is limited to the center of the string bed in a tennis racket so that frame vibrations are not excited by the impact of the ball. Therefore, a rigid body model was adopted for modeling the tennis racket frame. In this section, a uniform rigid rod representing the frame of a racket was used to derive the effective mass of the racket frame. Fig. 2 shows a ball and a rigid rod before and after impact between the ball and the rigid rod. The point mass, mb, represents the tennis ball, and the rod, mr, represents the tennis racket frame. R , rCMv , 2rv , 2rw , and RI are the distance from the center of mass of the rod to the impact point, the velocity at the center of mass of the rod, the velocity of the rod at the impact point, the angular velocity of the rod, and the moment of inertia of mass of the rod, respectively. The conservation of momentum is valid even for collision of bodies with energy loss during their collision"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003831_dscc2011-6060-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003831_dscc2011-6060-Figure1-1.png",
+        "caption": "Figure 1. DIAGRAM OF DUBINS VEHICLE AT POSITION ~rA THAT IS MOVING AT HEADING ANGLE \u03b8 AND TRACKING A RANDOMLYMOVING TARGET AT~rT WITH DISTANCE r = |~rA \u2212~rT | AND RELATIVE ANGLE \u03d5.",
+        "texts": [
+            " Next, the dynamic programming methodologies to compute the optimal control for this problem are provided, and the effectiveness of this approach is demonstrated for Brownian targets and targets with unknown trajectory. We conclude with a discussion and direction 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), and since we do not account for the possibility of antagonistic target trajectories, no knowledge of the UAV kinematics or state is assumed. 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 [14], bounded control u(t)\u2208 U \u2261{u : |u| \u2264 umax}, which has to be found. Note that by considering a model with more constraints (e.g., fixed vA and altitude, single integrator) tracking is more difficult for the UAV",
+            " In order for the control to be independent of the heading angle of the Dubins vehicle or the absolute position of the Dubins vehicle or target, we relate the problem to relative dynamics based on a time-varying coordinate system aligned with the direction of the Dubins vehicle velocity. The reduced system state is composed of the distance between the Dubins vehicle and target r = |~rT \u2212~rA| and the viewing angle \u03d5 between the Dubins vehicle\u2019s direction of motion and the vector from the Dubins vehicle to the target, as seen in Fig. 1: r = \u221a (\u2206x)2 +(\u2206y)2, \u03d5 = tan\u22121 ( \u2206y \u2206x ) . (3) The combined Dubins-target system (1-3) 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 infinitehorizon cost function W (\u00b7) with a distance-dependent discounting factor \u03b2(r) > 0 and with penalty \u03b5 (which may be zero) for 2 Copyright \u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/10/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use control: W (r,\u03d5,u) = Eu r    \u222b \u221e 0 e \u2212 t\u222b 0 \u03b2(r(s))ds k (r(t),u)dt    (4) k(r,u) = (r\u2212d)2 + \u03b5u2"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000038_iecon.2005.1569120-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000038_iecon.2005.1569120-Figure2-1.png",
+        "caption": "Fig. 2 Speed ripple effect chain.",
+        "texts": [
+            " Applying symmetrical voltages Vp=V and Vn=0, system (3) in steady-state conditions becomes: V 0 0 0             = Rs + j(xs + Xm ) j\u2206Xm / 2 jXm j\u2206Xm / 2 j\u2206Xm / 2 Rs 2s \u22121 + j(xs + Xm ) j\u2206Xm / 2 jXm jXm j\u2206Xm / 2 Rr s + jXr \u2206Rr s + j\u2206Xr / 2 j\u2206Xm / 2 jXm \u2206Rr s + j\u2206Xr / 2 Rr s + jX 'r                     I p In Irp Irn             (3\u2019) The positive sequence current Ip is the symmetrical current component at frequency f, while the negative sequence current In is the symmetrical current component at frequency (1-2s)f. The speed ripple is omitted but the In component computed directly by system (3\u2019) or indirectly by system (2\u2019) is still effective. In fact the speed ripple effect can be modeled as depicted in fig. 2. The stator negative component (I\u2019l) at frequency (1-2s)f produces a torque ripple \u2206T and consequently a speed ripple \u2206\u03c9r. This speed variation is seen by the stator as a mechanical angular variation that produces a phase modulation in the stator flux. Its series expansion shows that two new terms with the same amplitude and frequencies (1\u00b12s)f are added to the fundamental flux. These new fluxes induce e.m.fs El and Er with almost equal amplitudes at frequencies (1\u00b12s)f. El causes a left side component I\u201dl which decreases the previous component I\u2019l"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000899_978-3-540-88513-9_40-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000899_978-3-540-88513-9_40-Figure4-1.png",
+        "caption": "Fig. 4. The stepping motion of 3MLIPM in the frontal plane",
+        "texts": [
+            " The angle values of swing leg\u2019s actuators, which are used for pitching, can get from Fig. 3 by kinematic constrains just as (19). thighl and shinl are the length of thigh and shin. 2 2 2 0 2 2 3 4 1 2 2 0 2 2 2 0 3 0 2 1 2 arccos( ) 2 ( ) ( ) arccos( ) 2 arctan( ) 2 shin thigh shin thigh h sw h sw shin thigh shin h sw h sw l l l l l l x x z z l l l l l x x z z \u03b8 \u03b1 \u03b3 \u03b8 \u03b8 \u03b2 \u03b3 \u03b1 \u03b2 \u03c0 \u03b8 \u03b1 \u03c0\u03b3 \u03b3 \u03c0 \u03b3 + \u2212 = + = = + = \u2212 + \u2212 + \u2212 = = \u2212 \u2212 \u2212= + = \u2212 \u2212 . (19) The angle values of standing leg\u2019s actuators for pitching can be gotten in the same way. Fig. 4 shows the stepping motion of 3MLIPM in the frontal plane. With (5) and (6), (3) can be simplified to (20). 1 1uy vy w\u2212 = . (20) where 1(2 ) (2 ) ( ) 2 s s Lz u k v k w L k g \u03bb\u03b3 \u03bb= + = + = + . (21) With the initial conditions described in (22), it is easy to get the trajectory in y-axis of 1m as (23). 1 1 1(0) 0 ( ) 0 ( ) 0 2 T y y y T= = = . (22) 1 1 2 1 2 1 2 (1 ) ( ) , , 1 v Tv v vut t T u u u v T u w w e y t C e C e C C e C v v e \u2212 \u2212= + \u2212 = = \u2212 . (23) In the frontal plane, the two legs are kept parallel when the robot is walking. So it is simple to get the angle values of joints for rolling form Fig. 4. 1 1 5 1 6 1 10 1 1 arctan( ) 2 y z \u03c0\u03b8 \u03b8 \u03c0 \u03b8 \u03b8 \u03b8 \u03b8 \u03c0 \u03b8= + = \u2212 = = \u2212 . (24) The actual robot is 37 cm in height and 1.2 kg in weight. The shin and the thigh are both 5.2 cm in length. Fig. 5 and Table 1 gives an overview of the robot. Each leg of the robot has 5DOF. The hip joints (roll, pitch), the knee joint (pitch) and the ankle joints (roll, pitch) are actuated by servo motors. Each arm has 3DOF and the head has 1DOF, which are also driven by servo motors. The servo motors have drive circuits in themselves and can provide local position feedback control by themselves"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002691_icca.2010.5524420-Figure1-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002691_icca.2010.5524420-Figure1-1.png",
+        "caption": "Fig. 1 Four-force aircraft and tri-force aircraft",
+        "texts": [
+            ". INTRODUCTION Design and control of tri-rotor aircraft is a well-known problem and has attracted more attention in recent years. For a VTOL (vertically taking off and landing) aircraft, flying task can be realized by three forces (Fig. 1(b)). It is the minimum realization of controlling an aircraft with respect to four-force aircraft (Fig. 1(a)), for example, the Harrier. In practice tri-rotor aircraft is more valuable than fourrotor [1-7]. Some kinds of tri-rotor aircraft [8-11] were designed previously, at the same time, the corresponding controller was proposed to stabilize them. All those aircraft have the ability of hovering, however, they did not have the ability of high-speed forward flight. If fixed wings can be added into tri-rotor aircraft, then the aircraft not only have the ability of VTOL, but also have the ability of high-speed forward flight"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001644_978-90-481-2746-7_52-Figure5-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001644_978-90-481-2746-7_52-Figure5-1.png",
+        "caption": "Fig. 5. Equivalent stresses and displacements of the global model",
+        "texts": [
+            " The model was used to gain boundary conditions for more accurate models. On a basis of previous analyses [12,13] rivets weren\u2019t represented. Shell elements (Quad4, Tria3) and linear material models were used. Jerzi Kaniowski et al. 946 The boundary conditions were taken on a basis of the operational data. All moments were converted into forces and applied to the structure. Forces were applied to each rib, near spars. The model was fixed on one side. Nonlinear analysis was performed (Sol 106). Only geometrical nonlinearity was taken into account. Figure 5. presents equivalent stresses and displacements obtained in the analysis. On the upper skin there are some disturbances connected with local buckling, but there is no significant influence of this phenomenon on the lower skin. Fragment of the bottom skin, near rib no 21 (fig. 6) was chosen for local I level analysis. Methods for FEM analysis of riveted joints of thin walled aircraft structures 947 The Riveted joint FEM model was built for this region. The Presence of the rivet was taken into account, as well as the distance between middle surfaces of jointed parts"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002856_iecon.2011.6119562-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002856_iecon.2011.6119562-Figure4-1.png",
+        "caption": "Fig. 4 Results of the FE simulation. Left: all stator slot wedges existing. Right: one missing slot wedge.",
+        "texts": [
+            " Obviously the tip of the resulting phasor (gray arrow) moves along the dotted circle twice when the position of the missing slot wedge is changed over one electrical period. III. ANALYSIS OF TRANSIENT FLUX DISTRIBUTION BY SIMULATION In order to analyze the described influence of missing slot wedges on the current change and to support the explanations, a finite element (FE) simulation was done. Thus the change in the transient field path can be identified. The results of the FE simulations are depicted in Fig. 4. On the left side the healthy machine with all slot wedges placed, and on the right side a faulted case with one removed slot wedge are shown. The machine was operated without fundamental wave excitation. Only voltage pulses were applied in direction of phase axis U. The wedge was removed in a slot of phase winding U also (electrically orthogonal to the phase axis). As can be seen in the figure the path of the transient flux is changed from the symmetrical case (left) when removing the slot wedge (right)"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001580_1.2844956-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001580_1.2844956-Figure4-1.png",
+        "caption": "FIGURE 4. 1. Layout of Hermetically Sealed Container. 1 - Dewar Flask; 2 - Metal Rod; 3 - Holder Support; 4 - Electroacoustic Transducer (PZT); 5 - Gaskets (Foam Plastic); 6 - Holder Base (Ebonite); 7 - Cold Welding.",
+        "texts": [
+            " It is possible to show that in a classical approximation at temperatures higher than the Debyetemperature, the temperature dependence of body weight is described by the formula P = Mg, ^ 0 (6) where C is a the factor dependent on physical characteristics (including density and elasticity) of bodies and T is the absolute temperature. According to Equ. (6), an increase in the absolute body's temperature will cause a reduction of its weight. Such an effect was indeed observed in exact weighing of metal samples from nonmagnetic materials heated with ultrasound (Dmitriev, Nikushchenko and Snegov, 2003). The layout of the hermetically sealed container shown in Fig. 4. An example of the experimental dependence of a sample weight in the process of its heating and cooling is shown in Fig. 5. 2 0 ^ -2 I -6 I -8 5-10 (/3 ^-12 -14 -16 0 2 A ^^ 6 1 8 10 12 A A A A A A ' ' ^ A A \\ i ^ ^ A ^ ^ ' A A A ^ A A - FIGURE 5. Change in Mass of a Brass Rod. Ultrasound Frequency 131.25 kHz. The Touch Lines Indicate the Moments When the Ultrasound Was Switched On and Off. The temperature dependence of weight of various samples made of lead, copper, brass, titan and duralumin was measured"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003031_iros.2011.6094957-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003031_iros.2011.6094957-Figure3-1.png",
+        "caption": "Fig. 3. Steps in the computation of the graspability map: a) Different orientations for the same location of the hand base; b) Hand in one of the potential poses; c) Intersection of the workspaces \u03c6i with the object; d) Sets of reachable points \u03c8\u2032 i .",
+        "texts": [
+            " The steps in the algorithm are: Algorithm 1: Computation of the graspability map 1) Voxelize the parallelepiped delimiting the possible locations of the hand base frame around the object 2) Define a set \u0393 of potential locations and orientations for the hand base frame 3) For each pose of the hand base frame in \u0393 a) Check for collisions between the hand and the object. If there is a collision, discard the pose b) For each finger i compute \u03c8i = \u03c6i \u2229 \u2126 c) Obtain the sets \u03c8\u2032 i \u2282 \u03c8i of points with normals within the directions of force that each fingertip can apply d) If at least two sets \u03c8\u2032 i are not empty Verify the force closure condition Else Discard the pose 4) Return all the poses in \u0393 that lead to FC grasps Fig. 3 illustrates some steps in the computation of the graspability map for a banana, using a DLR hand II [22]. Fig. 3a shows different orientations for one potential location of the origin of the hand base frame. Fig. 3b shows the hand in one of these potential poses. Fig. 3c shows the intersection of the workspaces \u03c6i with the object, and Fig. 3d shows the corresponding sets \u03c8\u2032 i. A modified version of the Voxmap- Pointshell (VPS) algorithm [23] was used to compute the intersections, due to the fast responses (below 1 ms) to collision queries. This algorithm basically computes the intersections between voxmaps, voxelized volume structures for static objects, and pointshells, point clouds describing moving objects. For Step 3b, the computations are performed considering the workspaces \u03c6i as static objects, i.e. the global coordinate system is located in the hand base, and the coordinates of the points in \u2126 are transformed to that system via the transformation matrix describing the relative pose of the object with respect to the hand"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001319_j.ijpvp.2007.03.004-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001319_j.ijpvp.2007.03.004-Figure7-1.png",
+        "caption": "Fig. 7. Displacement of flange geometry 2 (FE solution) shows progressive loss of contact.",
+        "texts": [
+            " For these values of F 0 and K separation does not occur within the pressure range studied. Analysis is done for different internal pressure values up to 10MPa. The results predicted by the model are compared with nonlinear FE analysis in Fig. 6. The match is again seen to be good and the difference in y is about 1.6%, while the opening predicted from FEM is below half a micron (negligible for practical purposes even if it is reliable). Now we consider geometry 2 (see Fig. 1). This geometry has a different displacement characteristic as shown in Fig. 7. There is a gradual loss of contact between the two flange faces as the load increases. We study the effect of geometric nonlinearity in this case using FE analysis only. Nonlinear geometric effects can be included in ANSYS by giving the command NLGEOM,ON. Along with this we also consider the effect of stress stiffening (ANSYS command SSTIFF,ON) on the structure. We perform three different analyses for each set of loading parameters. (1) NLGEOM and SSTIFF ON (2) NLGEOM and SSTIFF OFF (3) NLGEOM ON and SSTIFF OFF"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002323_978-3-642-25486-4_25-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002323_978-3-642-25486-4_25-Figure3-1.png",
+        "caption": "Fig. 3. Visualization of relevant stiffness effects",
+        "texts": [
+            "= \u22c5 \u22c5 T P P,N diag P,NK J K J (18) As the definition of the Jacobean is altered the calculation is differing from the derivation in [7]. In our case the system\u2019s Jacobean is defined with the force transmission whereas it is commonly defined as velocity transmission matrix into the actuation space. Displacements of the end-effector caused by external wrenches can now be computed with the relationship: .= \u22c5-1 e P extx K F (19) The first six entries of the diagonal stiffness matrix Kdiag are corresponding to the extension stiffness of the forearm rods. Fig. 3 presents the resulting elasticity effects, K4 belongs to the forearms. K1 results from the bending stiffness of the upper arms about the Y-axis as well as the torsional stiffness of the gearbox unit. Both stiffness effects are located in a serial connection which is reducing the overall resulting stiffness K1 (about the Y-axis). Accordingly, K2 represents the elasticity effect resulting from the upper arm transverse bending stiffness about the Z-axis as well as the tilt stiffness of the gearbox unit"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000001_1.2032993-Figure3-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000001_1.2032993-Figure3-1.png",
+        "caption": "Fig. 3 Wedge action generate normal load",
+        "texts": [
+            " The amount of movement for a given pulling force is determined by the elastic support stiffness Ks of the elastic insert or clock spring. As the input shaft and sun roller rotate, the friction force Ff at the contact between the sun roller and loading planet drives the planet into the wedge gap towards the small end, wedging the loading planet firmly against the cylindrical raceways on the sun roller and outer ring. The wedge action generates appreciable normal forces N at the contacts between the loading planet and sun roller and between the loading planet and outer ring, as illustrated in Fig. 3. To balance the forces and maintain the position of the sun roller relative to the outer ring, two support planets are assembled beneath the sun roller in the wedge gap between the sun roller and the outer ring. Each support planet is radially supported through a bearing on a support shaft that is fixed to carrier housing. Torque capacity of the friction drive is determined to a large degree by the wedge action. For elastic materials, the normal force generated by wedging the loading planet in the wedge gap depends on structural compliances and local surface depression"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0003920_s11044-012-9320-0-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0003920_s11044-012-9320-0-Figure4-1.png",
+        "caption": "Fig. 4 Sketch of the ideal homokinetic joint",
+        "texts": [
+            " Homokinetic joints find many applications in mechanics and aerospace. They are specifically used in tiltrotor aircraft to allow the rotor to tilt with respect to the nacelle (cyclic flap motion in rotorcraft jargon) transmitting minimal tilting moment to the shaft. The above mentioned work contained in nuce the idea underlying the proposed vector formulation of kinematic joints, i.e., to prescribe the coincidence of the actual and imposed relative orientation of the two parts using the form of Eq. (8). As shown in Fig. 4, the relative orientation of the two parts is described in terms of three consecutive rotations, corresponding to a sequence of two universal joints whose initial and final angles, \u03d1 , are identical by construction, namely R(\u03d1,\u03d5) = exp(\u03d1e2 \u00d7 ) exp(\u03d5e1 \u00d7 ) exp(\u03d1e2 \u00d7 ); (42) \u03d1 and \u03d5/2 are the rotations about local axes 1 and 2 of each of the two universal joints, while the torque is transmitted about local axis 3. The orientation error \u03b5\u0302\u03b8 = ax ( exp\u22121 ( Ra abR(\u03d1,\u03d5)T )) = 0 (43) requires the relative orientation Ra ab to be compatible with the form described by Eq"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000318_isorc.2006.1-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000318_isorc.2006.1-Figure4-1.png",
+        "caption": "Figure 4. Event in an event area Wi.",
+        "texts": [],
+        "surrounding_texts": [
+            "For each event e occurring in an event area, multiple sensors gather QoE values of the event e. Then, each of the sensors sends sensed values to multiple actors. We assume every proper sensor sends the same sensed value for an event e to actors in an event area. Sensors might be arbitrarily faulty as presented before. Suppose there are li (> 1) sensors si1, \u00b7 \u00b7 \u00b7 , sili in an event area Wi. Sensors may be arbitrarily faulty. Here, we assume that at most fsi sensors out of li sensors are faulty in Wi. There are mi (> 1) actors ai1, \u00b7 \u00b7 \u00b7 , aimi in an event area Wi. Every actor may only suffer from stop fault. In this paper, we discuss how to realize WSAN tolerant of faulty of sensors. An actor has to first detect faulty sensors in an event area Wi. Each actor ais collects sensed values from multiple sensors si1, \u00b7 \u00b7 \u00b7 , sili and makes a decision on a value out of the values collected from the sensors. It is significant to first discuss how long an actor ais has to collect sensed values for an event. Let tik[e] show global time when a sensor sik senses an event e in an event area Wi. Let risk[e] denote global time when an actor ais receives a value of an event e from a sensor sik. Here, we make the following assumptions in each event area Wi: 1. For every pair of proper sensors sik and sih, |tik[e] - tih[e]| \u2264 \u03b1i for a constant \u03b1i if an event e occurs in an event area Wi. 2. |risk[e] - rish[e]| \u2264 \u03b2i for a constant \u03b2i if an actor ais receives sensed values of an event e from a pair of proper sensors sik and sih in Wi. Here, \u03b1i and \u03b2i are constants in an event area Wi. \u03b1i shows time difference between a pair of sensors which sense an event occurring in Wi. We assume at most one event occurs for every \u03b1i time units in Wi. The constant \u03b2i is given by the constant \u03b1i and the maximum difference \u03c4i of delay time from a sensor to an action in an event area Wi, i.e. \u03b2i = \u03b1i + 2\u03c4i. For example, \u03b1i, \u03b2i, and \u03c4i are in an order of \u03bc seconds. In a typical situation, we assume \u03b1i to be 40 [\u03bcsec] and \u03c4i to be 20 [\u03bcsec]. If |rish[e]\u2212risk[e]| \u2264 \u03b2i, an actor ais is referred to as simultaneously receive a sensed value e[sik] and e[sih] from a pair of sensors sik and sih, respectively. Hence, \u03b2i is minimum time difference to recognize a pair of different events. That is, if |rish[e] - risk[e\u2032]| > \u03b2i, the sensor as recognizes a pair of events e and e\u2032 to be different. Each actor receives sensed values from multiple sensors in the MAMS model. An actor ais collects sensed values from sensors for an event e. By using the values, an actor ais detects a sensor sik to be faulty in an event area Wi as follows: 1. If an actor ais does not receive any value from a sensor sik for \u03b2i time units since the actor ais received the first value from some sensor, the actor ais considers sik to be faulty. 2. An actor ais collects sensed values from sensors since ais receives the first message from some sensor. After it takes \u03b2i time units, the actor ais takes a majority value v in the values as a sensed value of an event, i.e. more than li/2 sensors send the value equivalent with v to ais [Figure 5]. Here, the actor ais considers a sensor sik to be faulty if sik sends a value v\u2032 different from the value v. If a pair of sensors sik and sih are proper, an actor ais simultaneously receives the sensed values e[sik] and e[sih]. Hence, e[sik] and e[sih] may be different. For example, a pair of temperature sensors take temperatures 15.1 and 15.2 [\u25e6C], respectively. An actor ais considers both the values to be equivelent and takes some value, e.g. the average value 15.15 [\u25e6C] of the values. Thus, an actor ais has to take one value v rom a collection V of the sensed values e[si1], \u00b7 \u00b7 \u00b7, e[sili] for the event e. 1. Calculate the median m in the value set V . 2. Remove a value v\u2032 such that |v\u2032 \u2212 m| > \u03b5i from V . 3. Calculate the median v of values in V . Proceedings of the Ninth IEEE International Symposium on Object and Component-Oriented Real-Time Distributed Computing 0-7695-2561-X/06 $20.00 \u00a9 2006 IEEE Here, if |e[sik] - v| \u2264 \u03b5i, a sensed value e[sik] is referred to as proper. A constant \u03b5i is decided by an application. If each actor ais receives sensed values from more than 2fsi + 1 sensors, at most fsi sensors are faulty and at least fsi + 1 sensors are proper. Hence, the actor ais takes the median value of the proper sensed values sent by the sensors. 4. Actors 4.1. Non-redundant execution of a method On receipt of sensed values of an event e from sensors, each actor ais makes a decision on what methods to be performed on what actuation device objects and then performs the methods on the actuation device objects in an event area Wi. In the multi-actor/multi-sensor (MAMS) model, each of multiple actors ai1, \u00b7 \u00b7 \u00b7 , aimi receives sensed values of an event e from multiple sensors si1, \u00b7 \u00b7 \u00b7, sili in the event area Wi. Suppose that a method op is decided to be performed on an object o by the actors in the event area Wi for an event e. If each actor ais makes a decision on the method op and performs the method op, the method op is mi(\u2265 1) times performed on the object o. Here, the state of the object o may get inconsistent. For example, suppose each of two actors ais and ait receives sensed values of an event e on the temperature. Each of the actors ais and ait makes a decision on warming the air with 2\u25e6C degree. Then, each of the actors ais and ait issues a method up(2) to the air-conditioner object ac. The temperature is in result increased by 4\u25e6C because a pair of up(2) methods are performed on the object ac. Here, the method up(2) should be performed only once on the air-conditioner object ac for the event e even if multiple actors receive the sensed information of the event e from sensors. In the MAMS model, the redundant invocations of a method by multiple actors have to be resolved. Next, suppose each of the actors ais and ait issues a method temp to an air-conditioner object ac. Here, the state of the air-conditioner object ac is not changed. Thus, even if multiple actors multiple times issue an idempotent method like temp to an object, the state of the object does not get inconsistent. Following the examples, we classify methods supported by objects into the following types of methods with respect to whether or not the state of an object is changed. 1. Change method. 2. Non-change method. Through a change type of a method op, state of an object in an event area is changed, i.e. op(s) = s for some state s. The method up is a change type on an air-conditioner object ac. On the other hand, the state of an object is not changed by a non-change method, i.e. op(s) = s for every state s. The method temp is a non-change method. For each event e, a change type method op can be performed only once even if multiple instances of the method op are issued to the object by multiple actors. A non-change method can be performed multiple times on an object, i.e. idempotent. However, the performance of the actuation device object may be degraded if more number of methods are performed on the object. There are following approaches to realizing the unique execution of a method on an actuation device object: 1. Actor-side approach. 2. Object-side approach. In the actor-side approach, only one method is issued to an object from multiple actors. On receipt of a method, the method is just performed on an object. In one way, the actors cooperate with each other to make a consensus on which actor to issue a method to an object o. Then, only the selected actor issues the method while the other actors do not issue the method to the object o. It takes time to exchange messages among the actors, e.g. takes three rounds if the two-phase commitment protocol [18] is taken. In the object-side approach, an actuation device object o takes a method op only once even if each of multiple actors sends the method to the object o. Each actor issues a method to an object independently of other actors. Each of the actors issues a method op to an object o. Here, each instance of the method op to be taken for an event should be uniquely identified. On receipt of a method op from an actor, the identifier id of an instance of the method op is checked on an object o. If the method with the identifier id had not yet been performed on the object o, the method op is performed on the object o. Then, the identifier id is recorded in the log. If the identifier id of the method op is found in the log, the method op is not performed since another instance of the method op has been already performed on the object o. In WSAN, realtime communication is in nature required to deliver sensed values to actors Proceedings of the Ninth IEEE International Symposium on Object and Component-Oriented Real-Time Distributed Computing 0-7695-2561-X/06 $20.00 \u00a9 2006 IEEE and methods to actuation device objects. Hence, we take the object-side approach by giving the unique identifier to each method since no communication among the actors is required. Each instance opis of a method op issued by an actor ais is identified in a pair of method type op and global time tis when the actor ais receives an event. Here, a pair of identifiers \u3008op, tis\u3009 and \u3008op\u2032, tit\u3009 of instances opis and opit, respectively, are referred to as temporarily equivalent (opis \u223c= opit) iff op = op\u2032 and |tis \u2212 tit| \u2264 \u03b1i. A method opit is considered to be redundant or an object o, i.e. opit is rejected if a method opis had been performed on the object o such that opis \u223c= opit. On each actuation device object, each method can be only once performed even if multiple actors send instances of the method to the object."
+        ]
+    },
+    {
+        "image_filename": "designv11_25_0001952_ichr.2009.5379580-Figure7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001952_ichr.2009.5379580-Figure7-1.png",
+        "caption": "Fig. 7. Computing temporary goals.",
+        "texts": [
+            " These two goals describe the maximal reachable configurations on the path when considering the two parameters vc, the maximal allowed motor velocity and vw , the maximal workspace velocity for a joint. It is possible to define different Vc and Vw values for each joint, but in general the same value is used for all joints. When computing the goals, the maximal allowed position in the future is computed, so that the speed of no joint exceeds the maximal allowed velocities in C Space and in workspace. In Fig. 7 the search for a new goal starts at C~obot and both goal limits gc and gw are calculated. Although Vc would allow robot to move until gc, this would violate the workspace velocity limit V w which means that there would be a joint moving too fast. So g w will become the next temporary goal gnext. The direction of further movement in C-Space is given by v = gnext - Cro bot , and in case v does not violate C-Space or workspace speed limits, it can be passed to the low level controllers. In Fig. 8 the results of two simulation experiments are shown"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0002810_20110828-6-it-1002.01052-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0002810_20110828-6-it-1002.01052-Figure2-1.png",
+        "caption": "Fig. 2. Throttle Body scheme",
+        "texts": [
+            " As a consequence, ECU varies the amount of injected fuel in order to maintain the desired air-fuel ratio. The reference signal is the solution of a trade-off between the driver request (acceleration pedal position) and the effective traction possibilities depending upon driveability, safety and emission constraints. The control signal generated by the ECU becomes, by means of an H-bridge power converter, the armature voltage of a dcmotor. The rotation motion is then transferred from the motor shaft to the plate shaft through a gear system. A schematic of the ETB is shown in Fig. 2. Despite its apparent simplicity, the system behavior is affected by many nonlinearities which can dramatically alter its dynamics. They can be briefly summarized as follows. \u2022 Piece-Wise Linear Restoring Torque. When a failure of the dc motor occurs, for safety reasons it is necessary to ensure that the valve comes back to a default position (called limp-home position) (Vasak et al., 2003). To guarantee the limp-home, two additional springs are used. The resulting elastic torque is a piece-wise linear function of all the admissible angles"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001683_asjc.272-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001683_asjc.272-Figure2-1.png",
+        "caption": "Fig. 2. The experimental setup.",
+        "texts": [
+            " The advantage of a controller implemented by FPGA includes shorter development cycles, lower cost, small size, fast system execute speed, and high flexibility [28]. The Quartus II software is the development tool for programmable logic devices. The Nios II processor is a configurable, versatile, RISC embedded processor. It can be embedded into Altera FPGA, and allow designers to integrate peripheral circuits and processors in the same chip. Additionally, the PC-developed algorithm and C language program can be rapidly migrated to the Nios II processor to shorten the system development cycle [28]. The experimental setup is shown in Fig. 2. This study used the Altera Stratix II series FPGA chip, and the Altera Quartus II software (Altera, City, Country), and the verilog hardware description language was used to implement the hardware control system. The proposed control algorithm is realized in the Nios II programming interface. The software flowchart of the control algorithm is shown in Fig. 3. In the main program, the initialization of controller parameters is preceded. Next, the interruption interval for the interruption service routine (ISR) with a 1msec sampling rate is set"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000379_9780470058480-Figure1.7-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000379_9780470058480-Figure1.7-1.png",
+        "caption": "Figure 1.7 Force summing bellows arrangement",
+        "texts": [
+            " V1 R + V2 R + V3 R =\u2212 ( V0 R ) The negative output is inherent in the hardware since the current flow in the feedback resistor is away from the summing junction and hence V0 must be negative. Thus we have a simple summing amplifier. It is also possible to select different impedances for each input as well as the feedback to generate complex dynamic transfer functions. Force Summing Bellows Another commonly used summing device uses bellows and a summing link to generate a displacement proportional to two (or more) pressures in both hydraulic and pneumatic systems. Figure 1.7 shows such a device used to modulate the opening of a flapper valve as part of a servo mechanism. This arrangement is ideal since the servo flapper displacement is extremely small and therefore any error induced by the spring rate of the bellows will be negligible thus providing accurate force summing. Differential Equations 11 This section reviews the differential equation which is the standard mathematical approach to defining the dynamic behavior of physical systems. Here we will limit our attention to linear differential equations with constant coefficients leaving how to deal with nonlinearities until later in the book"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0001699_s1068798x10100072-Figure4-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0001699_s1068798x10100072-Figure4-1.png",
+        "caption": "Fig. 4. Determining the radius of curvature of a cycloid crown profile.",
+        "texts": [
+            " To this end, we write two vector equations + = + = If we denote the vector moduli by e, lMP, rw1, respectively, we may write these equations as expansions in terms of the unit vectors running along the axes of motionless coordinate system S. We obtain two equations (6) (7) Simultaneous solution of Eqs. (6) and (7) yields the following formulas in terms of the initial parameters of the transmission (8) (9) Thus, the radius of curvature of the cycloid cam profile is (10) To determine the radius of curvature of the cycloid crown profile, we employ a construction analogous to that in the previous case, but for engagement of the ring and crown (Fig. 4). Consider the roller contact with the internal cycloid profile at point K '. In this case, the radius of the generating circle rw2 remains constant, as does the radius of the circle rcy corre sponding to the centers. Thus, Eqs. (1) and (2) may lOrP rw2 1 \u03c72 2\u03c7 \u03d5cos\u2013+ ;= \u03b1sin \u03c7rw2 \u03d5sin /lOrP;= \u03b1cos 1 \u03c7 \u03d5cos\u2013( )rw2/lOrP;= \u03b1tan \u03c7 \u03d5sin / 1 \u03c7 \u03d5cos\u2013( );= lO2C rw2/ \u03d5sin \u03b1tan \u03d5cos\u2013( ).= \u23a9 \u23aa \u23aa \u23aa \u23aa \u23a8 \u23aa \u23aa \u23aa \u23aa \u23a7 O2C CO1 O1O2; O1M MP O1P. CO1, O2O1, O1M, MP, O1P, lCO1 , lO1M, (i, j) lO2C \u03d5isin \u03d5jcos\u2013( ) lCO1 \u03b2isin \u03b2jcos+( )\u2013 e;= lO1M \u03b2isin \u03b2jcos\u2013( )\u2013 lMP \u03b1isin \u03b1jcos\u2013( )+ rw1"
+        ],
+        "surrounding_texts": []
+    },
+    {
+        "image_filename": "designv11_25_0000699_wcica.2008.4593160-Figure2-1.png",
+        "original_path": "designv11-25/openalex_figure/designv11_25_0000699_wcica.2008.4593160-Figure2-1.png",
+        "caption": "Fig. 2. 6-axis hybrid vibration isolation system based on Stewart platform with cubic configuration: (a) the perspective view and (b) the cross-section view.",
+        "texts": [
+            " 1613 developed in this paper, a force feedback control principle to isolate the displacement vibrations on the lower platform is proposed and analyzed. According to the developed dynamic model and control method, the numerical simulation results on the vibration isolation characteristics of the developed vibration isolation platform are presented. II. -AXIS HYBRID VIBRATION ISOLATION SYSTEM BASED ON CUBIC CONFIGURATION OF STEWART PLATFORM The developed 6-axis hybrid vibration isolation system based on cubic configuration of Stewart platform is shown in Fig. 2, in which Figs. 2(a) and 2(b) show the perspective view and the cross-section view, respectively. According to Fig. 2, the upper and lower platforms are connected together by six hybrid struts and each hybrid strut is connected to the upper and lower platforms by the 3-DOF flexible joint, respectively. Each hybrid strut consists of a hybrid actuator, a force sensor, two flexible joints, and a link rod. Each hybrid actuator consists of an actuator, a spring, and a damper. The flexible joints not only possess the characteristics of the general spherical joint but also can eliminate the friction and backlash. The upper platform is a disk with internal and external radii of 1R and 2R , respectively. Through structural design, the center of gravity (C.G.) of the upper platform positions at the geometric center K of the hexapod as shown in Fig. 1. The lower platform also has a configuration of a disk with the appropriate diameter. Fig. 2(b) shows the cross-section view of the 6-axis hybrid vibration isolation system, from which it can be seen that the arbitrary two struts are orthogonal and the length L of each strut is the distance between the two intersections of the strut axis with the two adjacent strut axes. To model the dynamics of the 6-axis hybrid vibration isolation platform, three coordinate systems are established as shown in Figs. 1 and 2. The universal reference frame B , which has its origin at the C.G. of the Stewart platform, and the two mobile frames P and R affixed on the upper and lower platforms, respectively",
+            " In the dynamic modeling of the 6-axis hybrid vibration isolation platform, two assumptions are used, which include A, the vibration amplitudes of the upper and lower platforms are small, in this case, the elongation of each strut is also small and the kinematic configuration of the Stewart platform remains almost unchanged and B, the interaction forces between the upper platform and the struts are in the axial directions of the corresponding struts because of the flexible joint. III. DYNAMIC MODELING Considering the 6-axis hybrid vibration isolation system shown in Fig. 2, the relationship between the elongation velocities of the six struts and the velocity vector of upper platform can be expressed as = v Jq (1) where q is the 16\u00d7 vector of axial velocities of the ends of the six struts in the upper platform; J is the 66\u00d7 Jacobian matrix of the upper platform; v is the 13\u00d7 vector of velocities of the origin in the mobile frame P ; is the 13\u00d7 vector of angular velocity of P in B . As shown in Fig. 1, the Jacobian matrix of the upper platform in the reference frame B can be expressed as \u2212\u2212\u2212 \u2212\u2212 \u2212 \u2212 \u2212\u2212\u2212 \u2212\u2212 = 20202 2232231 2232231 20202 2232231 2232231 6 1 LL LLL LLL LL LLL LLL J (2) where L is the length of struts"
+        ],
+        "surrounding_texts": []
+    }
+]
\ No newline at end of file