多功能嬰兒護理車的設(shè)計【全新多功能嬰兒車】【說明書+CAD+SOLIDWORKS+STP】

多功能嬰兒護理車的設(shè)計【全新多功能嬰兒車】【說明書+CAD+SOLIDWORKS+STP】,全新多功能嬰兒車,說明書+CAD+SOLIDWORKS+STP,多功能嬰兒護理車的設(shè)計【全新多功能嬰兒車】【說明書+CAD+SOLIDWORKS+STP】,多功能,嬰兒,護理,設(shè)計,全新,說明書,仿單 TECHNICAL PAPER A method of reducing the windage power loss in a laser scanner motor using spiral-groove aerodynamic thrust bearings functioning as a viscous vacuum pump Shigeka Yoshimoto Masaaki Miyatake Tomoatsu Iwasa Akiyoshi Takahashi Received: 29 June 2006 / Accepted: 2 November 2006 / Published online: 1 December 2006 C211 Springer-Verlag 2006 Abstract We propose a spiral-groove aerodynamic thrust bearing functioning as a viscous vacuum pump in a laser scanner motor to reduce the windage power loss of a polygon mirror. The proposed bearing pumps out the air in the scanner housing using the pumping effect of the spiral-groove thrust bearing, reducing the inner pressure of the housing. The pumping performances and the static characteristics of the spiral-groove thrust bearings were investigated numerically and experi- mentally. Two numerical calculation methods were used to study the pumping characteristics of the spiral- groove thrust bearing. It was found that a bearing with 15 spiral grooves reduced the inner pressure of the housing to 0.01 in the bearing clearance, and that the slip flow in the bearing clearance influenced the pumping characteristics of the viscous vacuum pump. As mentioned above, we proposed using a spiral- groove aerodynamic thrust as a viscous vacuum pump, which would enable us to reduce the windage power loss of the polygon mirror in a laser scanner motor, and reduce the size of the laser scanner motor. In addition, the pumping and load-carrying characteristics of our proposed bearing were investigated theoretically and experimentally. Two numerical calculation methods considering the slip flow were used. The first method employed was the narrow-groove theory reported by Whipple (1958), Vohr and Chow (1965) and Malanoski and Pan (1965), and the second method used was the divergence formulation (DF) method using the boundary-fitted coordinate system proposed by Ka- wabata (1986). Therefore, the objective of this work was to confirm the usefulness of our proposed bearing for a laser scanner motor, both theoretically and experimentally. 2 The proposed scanner motor using a spiral-groove aerodynamic thrust bearing as a viscous vacuum pump Figure 3 shows the structure of a laser scanner motor using spiral-groove aerodynamic thrust bearings func- tioning as a viscous vacuum pump. The shaft is fixed to the housing, and the rotor and mirror rotate. Two identical aerodynamic thrust bearings with spiral grooves are located at the upper and lower parts of the rotor. In the inner circle of the thrust bearing, several small holes are drilled so that the air inside the housing can be pumped out through these holes. The housing is sealed tightly to maintain a high vacuum in the hous- ing. Figure 4 shows the geometrical configuration of the thrust bearing with spiral grooves, along with the symbols used in this paper. A number of spiral grooves with groove tilt angle, b, are formed at even intervals in the circumferential direction on the bearing surface. In addition, a land region, with length, l w (where l w = r 1 r 2 ), is created, allowing the bearing to both pump air and support the weight of the rotor and mirror. 3 Numerical calculation methods Two calculation methods were used to obtain the pumping characteristics of the proposed thrust bearing numerically. The first method used was narrow-groove theory, and the second method was the DF method with a boundary-fitted coordinate system. Further- more, the first-order slip flow proposed by Burgdorfer (1959) was considered in the bearing clearance in both calculations. 3.1 Narrow-groove theory In narrow-groove theory presented by Malanoski and Pan (1965), it is assumed that the number of spiral grooves is infinite, and that the width is infinitesimal. Since there is no pressure gradient in the circumfer- ential direction, the mass flow rate in the spiral-groove region in the r-direction is q C0 p RT k 1 dp dr C0 k 2 r cos b C26C27 rdh; 1 where k 0 1 C0 aA g aA r k 1 A g A r a1 C0 aA g C0 A r 2 sin 2 b no. 12lk 0 k 2 xa1 C0 ah g C0 h r A g C0 A r sin b C14 2k 0 and A g h 3 g 1 6k h g C18C19 ; A r h 3 r 1 6k h r C18C19 : 2 For the land region in the inner circle of the thrust bearing, the mass flow rate is q C0 p 12lRT h 3 r 1 6k h r C18C19 dp dr rdh: 3 To obtain the distribution in pressure in numerical calculations, the equation of continuity is assumed in a small element in the bearing clearance using Eqs. (1) and (3). The derived equation is solved numerically assuming the following boundary conditions, at Microsyst Technol (2007) 13:11231130 1125 123 r r 2 ; p p a ; and at r r 0 ; p p u ; 4 where p u is defined as the ultimate pressure where no flow rate exists in the bearing clearance. Therefore, the ultimate pressure, p u , can be obtained by solving Eq. (1) under the boundary condition where q =0. Furthermore, when a load is imposed on the rotor, the continuity of mass flow rate is assumed between the upper and lower bearing clearances, and therefore the inner housing pressure can be obtained. 3.2 DF method with boundary-fitted coordinate system In narrow-groove theory, the number of grooves is assumed to be infinite and hence, the effect of the number of grooves on the pumping characteristics is not applicable using this theory. Therefore, we used the DF method with a boundary-fitted coordinate system to clarify the effect of the number of grooves. The mass flow rates in the boundary-fitted coordinate system, n g coordinates, as shown in Fig. 5, were derived by transformation from the polar r h coor- dinates as presented by Kawabata (1986). For the mass flow rate in the n-direction q n p RT a p C0A p n B p g D C18C19 : 5 For the mass flow rate in the g-direction q g p RT c p B p n C0 C p g E C18C19 6 where A h 3 12l 1 6k h C18C19 a J B h 3 12l 1 6k h C18C19 b J C h 3 12l 1 6k h C18C19 c J D C0 rx 2 h r g E rx 2 h r n a r g C18C19 2 rh g C18C19 2 b r n r g rh n rh g c r n C18C19 2 rh n C18C19 2 J r n rh g C0 r g rh n : Figure 5 shows a control volume in the bearing clearance where a continuity of the mass flow rate was assumed. Therefore, the mass flow rate in the normal direction to the constant n or g boundary is given by Q n Z g 2 g 1 p RT C0A p n B p g D C18C19 dg Q g Z n 2 n 1 p RT B p n C0 C p g E C18C19 dn: 7 The following equation is derived from the continuity of mass flow rate Rotor Pump out Sealed housing Polygon mirror Motor Pump out Exhaust holes Thrust bearing with spiral grooves Journal bearing Fig. 3 A laser scanner motor using spiral-groove aerodynamic thrust bearings as a viscous vacuum pump Groove Ridge Shaft Rotor Pump in Pump in Groove e r 0 r 1 r 2 hd Groove Ridge Land Thrust bearing Exhaust holes h r h d Pump out r hr 0 region 1- a Fig. 4 The configuration of the spiral grooves and symbols 1126 Microsyst Technol (2007) 13:11231130 123 Q n 2A1 Q n 1A3 C0 Q n 2A2 C0 Q n 1A4 Q g 2A1 Q g 1A2 C0 Q g 2A3 C0 Q g 1A4 0: 8 3.3 Calculation of the pressure in a sealed housing with time The pressure in the sealed housing was calculated using Eq. (1) after the rotor had begun to rotate. Assuming that no axial load was imposed on the rotor, and that the upper and lower bearing clearances had the same value, then, in the narrow-groove theory, the mass flow rate in unit time is given by Qt2 Z 2p 0 qj rr2 r 2 dh: 9 Next, assuming that the mass of air in the housing with a volume, V i, was G(t) at time t, then, under isothermal conditions, G(t + Dt) is expressed as Gt DtGtC0QtDt: 10 Furthermore, using pressure, p i , and density, q i , in the housing, the mass of air included in the housing at time t, is given by Gtq i tV i p i tV i =RT *q i tp i t=RT : 11 Substituting Eq. (11) into Eq. (10), the housing pressure at time, t+Dt, is given as follows. p i t Dtp i tC0QtRTDt=V i : 12 In our calculations, Dt = 0.0167 s, and at t = 0, the initial value of the housing pressure is assumed to be atmospheric pressure. The housing pressure obtained by Eq. (12) is used as the boundary condition of Eq. (1) for the next time step. 4 Results of the calculations As mentioned above, narrow-groove theory cannot be used to estimate the effect of the number of grooves on the pumping characteristics of the spiral-groove bear- ing, although this theory is very simple, and is very useful for designing aerodynamic bearings with her- ringbone or spiral grooves. Therefore, the suitability of narrow-groove theory to predict the pumping charac- teristics of the spiral-groove thrust bearing treated in this paper was confirmed by comparing data obtained using the DF method. The principal dimensions used in our calculations are shown in Table 1. Figure 6 shows the pressure distribution in the bearing clearances obtained using the DF method with a boundary-fitted coordinate system for n = 8 and 15. The DF method can be used to calculate the pressure distribution by considering the groove shape, as shown in Fig. 6. Accordingly, it can be seen that the pressure distribution varied between the grooves in the cir- cumferential direction. By increasing the number of grooves, the pressure variation between the grooves decreased, and it can be predicted that the pressure distribution would approach that obtained using nar- row-groove theory. Figure 7 shows a comparison of the theoretical ultimate pressure obtained using narrow-groove theory and the DF method considering the slip flow in the bearing clearance. The number of grooves was changed from 6 to 15 using the DF method. In Fig. 7, a theo- retical result without considering the slip-flow effect using narrow-groove theory is also shown for refer- ence. As can be seen in Fig. 7, the ultimate pressure for n = 15 using the DF method had almost the same value as that using narrow-groove theory, and the ultimate pressure increased with decreasing the number of grooves. However, the effect of the number of grooves was not large under the conditions shown in Fig. 7. From these observations, the theoretical calculations discussed below were carried out using narrow-groove theory because the spiral-groove bearing with n =15 was used in the experiment. A1 A3 A4 A2 i i i j j j Q 2A1 Q 1A2 Q 2A3 Q 1A4 Q 2A1 Q 1A3 Q 2A2 Q 1A4 r q x x x x x x h h h h h h Fig. 5 The n and g coordinates and mass flow continuity in a control volume Table 1 Principal dimensions of the spiral-groove thrust bearing r 0 (mm) r 1 (mm) r 2 (mm) ab(C176) h d (lm) n 14.0 9.0 8.5 0.5 15.0 12.0 15 Microsyst Technol (2007) 13:11231130 1127 123 Figure 8 shows the theoretical pressure distribution with and without the inner land region for a dimen- sionless axial displacement of 0.4 to clarify the effect of the inner land region on the load capacity. When there was no inner land region, the difference between the pressure distributions in the upper and lower bearing clearances was very small, and in this case even a negative load capacity was generated. In contrast, the difference became large in the bearings with an inner land region. In addition, the ultimate pressure, P u , was not sensitive to the presence of an inner land region. Figure 9 shows the relationship between the dimensionless axial displacement and the dimension- less load capacity at 20,000 rpm. In Fig. 9, the load capacity of the proposed bearing is compared to that of a conventional bearing that was operated at atmo- spheric pressure. The load capacity of the proposed Fig. 6 Pressure distribution obtained using the DF method and spiral-groove shapes 0 10000 20000 30000 0 0.2 0.4 0.6 0.8 1 Rotational Speed : N rpm Ultimate Pressure : P u a=0.5 b = 15deg. h r0 =2.5 m DF Method (Slip Flow) n= 6 8 15 Narrow Groove Theory No Slip Flow Slip Flow Fig. 7 Effect of the number of spiral grooves on the ultimate pressure in the housing 0.2 0.4 0.6 0.8 1 0.5 1 1.5 0 Dimensionless Pressure P R Coordinate Land Region Groove Region : With Land Region l w =0.5mm : No Land Region h r0 =4m : 30000rpm :=0.4 Lower Bearing Pressure Upper Bearing Pressure Fig. 8 Effect of the inner land region on the pressure distribu- tion 1128 Microsyst Technol (2007) 13:11231130 123 bearing showed relatively smaller values compared to those of the conventional bearing, but it was large enough to support the scanner rotor. The dimension- less load capacity of 0.1 corresponds to a load of about 4N in the proposed bearing. 5 Comparison with the experimental results To verify the calculated results, we conducted a series of experiments. Figure 10 shows the experimental apparatus used to measure the pumping characteristics and the load capacity of the proposed bearing. The polygon mirror was attached to one end of the rotor, which was supported by three aerostatic journal bear- ings. The polygon mirror was located between two identical aerodynamic thrust bearings with spiral grooves. The mirror and the thrust bearings were made of ceramic. The outer space of the polygon mirror was connected by a tube to a tightly sealed housing with a volume of 65 cm 3 . The pressure inside the housing was measured using a vacuum pressure gauge. The bearing clearance of the thrust bearing could be changed using spacers with different thicknesses, denoted by Plate A in Fig. 10. The rotor was operated using a DC motor. Figure 11 shows the relationship between the inner pressure of the housing and the elapsed time after the rotor had begun rotating. The inner pressure decreased rapidly immediately after the rotor had begun rotating, and leveled out within a period of 5 min. 0. 2 0. 4 0. 6 0.8 0.2 0.4 0.6 0.8 0 Dimensionless Axial Displacement Dimensionless load-carrying capacity W 20000rpm Present Conv. h r0 =3.0m : h r0 =4.0m : h r0 =5.0m : Fig. 9 The relationship between the axial displacement and the load capacity Tube Digital Tachometer Polygon Mirror Plate A Motor Housing Aerostatic Journal Bearings Shaft Spiral-Grooved Thrust Bearing Pressure Gauge Volume Air out Air out Air in Fig. 10 The experimental apparatus 10 20 0.2 0.4 0.6 0.8 1 0 Cal. Exp. : h r0 =2.5 m : h r0 =3.5 m : h r0 =4.5 m Elapsed time : t min Dimensionless Pressure in the Chamber P i l w =0.5mm : 20000rpm P u Fig. 11 Variation of the housing pressure with elapsed time 10000 20000 30000 0.2 0.4 0.6 0.8 1 0 Rotational speed N rpm Ultimae Pressure P u Exp. Cal. : h r0 =2.5m : h r0 =3.5m : h r0 =4.5m Fig. 12 Effect of the rotational speed and bearing clearance on the ultimate pressure Microsyst Technol (2007) 13:11231130 1129 123 Figure 12 shows the ultimate pressure when the rotational speed and the bearing clearance were changed. As can be seen in Fig. 12, the proposed thrust bearing with spiral grooves decreased the inner pres- sure of the housing to 0.01 MPa at h r0 = 2.5 lmata speed of 20,000 rpm. Figure 13 shows the variation in the ultimate pres- sure in the housing when a load was imposed on the bearings. In this experiment, a rotor mass = 180 g was used to impose a load on the bearings. The test rig was set on a magnetic chuck, which could be tilted from the horizontal to a vertical angle. As can be seen in Fig. 13, the ultimate pressure was not greatly influenced by the imposed load, and the proposed bearing could support the rotor without any deterioration in the pumping characteristics. 6 Conclusions We have developed a spiral-groove aerodynamic thrust bearing functioning as a vacuum pump to reduce the windage power loss of the polygon mirror and the size of a laser scanner motor. The pumping characteristics of the proposed bearing were investigated theoretically and experimentally. As a result, the following conclu- sions were derived: 1. The proposed bearing with spiral grooves reduced the pressure in the sealed housing to 0.01 MPa at h r0 = 2.5 lm at 20,000 rpm. 2. Though the load capacity of the proposed bearing was relatively small compared with that of a bearing operating at atmospheric pressure, it was large enough to support the rotor mass in this type of device. 3. Calculation method presented in this paper can predict well the pumping characteristics of the proposed bearing. References Burgdorfer A (1959) The influence of the molecular mean free path on the performance of hydrodynamic gas lubricated bearings. Trans ASME J Basic Eng 81(1):94100 James DD, Potter AF (1967) Numerical analysis of the gas- lubricated spiral-groove thrust bearing-compressor. Trans ASME J Lub Tech 89(4):439444 Kawabata N (1986) A study on the numerical analysis of fluid film lubrication by the boundary-fitted coordinates system (first report, fundamental equations of df method and the case of incompressible lubrication). Trans JSME Ser C 53(494):21552160 Sato Y, Ono K, Iwama A (1990) The optimum groove geometry for spiral groove viscous pumps. 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Trans ASME J Trib 122(1):131136 Malanoski SB, Pan CHT (1965) The static and dynamic characteristics of the spiral-grooved thrust bearing. Trans ASME J Basic Eng 87:547558 1 2 0.2 0.4 0.6 0.8 1 0 Imposed Load w N Ultimae Pressure P u 20000rpm Exp. Cal. : h r0 =2.5m : h r0 =3.5m : h r0 =4.5m Fig. 13 Effect of the imposed load on the ultimate pressure 1130 Microsyst Technol (2007) 13:11231130 123