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【机械类毕业论文中英文对照文献翻译】采用螺旋槽推力轴承的电动机作为一种粘性的运作真空泵降低激光扫描仪功率损耗的一种理论
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机械类毕业论文中英文对照文献翻译
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【机械类毕业论文中英文对照文献翻译】采用螺旋槽推力轴承的电动机作为一种粘性的运作真空泵降低激光扫描仪功率损耗的一种理论,机械类毕业论文中英文对照文献翻译,机械类,毕业论文,中英文,对照,文献,翻译,采用,螺旋,推力,轴承,电动机,作为,一种,粘性,运作,真空泵,降低,激光,扫描仪,功率,损耗,理论
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采用螺旋槽推力轴承的电动机作为一种粘性的运作真空泵降低激光扫描仪功率损耗的一种理论摘要 我们提出了一种采用螺旋槽推力轴承的电动机作为一种粘性的运作真空泵降低激光扫描仪风阻的功率损耗。使用螺旋槽推力轴承扫描仪在空气中抽吸理论的提出，减少了内压力的空间。抽水的表现和螺旋槽推力轴承的静态特性是经过数值模拟和实验调查过的。这两个数值的计算理论被用于研究螺旋槽推力轴承的抽水特性。结果发现：一个具有15螺旋沟槽的轴承在2000转和2.5微米的轴承间隙使使空间的压力减少到了0.01MPa以下，这已经足够作为支持旋转轴的承载能力。1、简介 激光扫描马达转速近年来显著增长，但是由于旋转速度的增加，激光多能镜扫描仪引起的功率损耗也急剧增加。这个速度增加也导致其他问题，比如温度的显著上升和电机功率消耗的显著增加。因此在扫描器电机的外壳内充满氦气是减少风阻功率消耗的有效措施。但是这种方法也有不好的地方，要长时间限制空间温度上升是非常困难的。 吉本和高桥（1999）提出了一个新的方法来降低多棱镜的风阻功率损耗。他们的方法是使用一个人字形气体轴颈轴承作为一个真空泵来进行抽吸如图1.当电动机开始旋转，由于真空泵的粘性效果，空间中的气体被抽吸了出来。据报道：在这个多棱镜激光扫描仪风阻电机功率损耗是常规电机转速达到30000转时的功率消耗的一半如图2。 顺便说一下，小型化的影印机和激光打印机要有一个更高的输出速度这种方式也是需要的。因此这些设备小型化是目前需要解决的重要问题。激光扫描马达所示图一：被用作一人字形沟槽真空泵。因此，这一扫描仪的小型化是通过缩短泵的长度来防止泵的性能恶化。因此，我们建议用一个螺旋槽推力轴承的真空泵更容易实现与真空泵电机的小型化激光扫描仪.，支持扫描仪镜和转子重量的建议也被提了出来。要确立我们提出的优化设计轴承，螺旋槽推力轴承的抽水特点也不得不在论上进行确定。一些研究人员对压缩机螺旋槽轴承的抽水特点进行了研究，詹姆斯和波特（1967）对螺旋槽推力轴承用作压缩机的性能的数值进行了分析。佐藤等人（1990年）和佐藤和骑士（1992）报告了人字形和螺旋槽粘性泵的性能，并讨论了压缩机的最佳槽的几何形状，田中村木（1991）开发了一种激光扫描马达，该扫描仪的主轴是由一个螺旋槽径向轴承产生的压缩空气来支持的。吉本等（2000年）的理论和实验研究了气体在粘性真空滑动轴承中被抽吸的现象。他们指出大于0.01的轴承间隙Knudsen数影响了粘性真空泵的抽水特点。如上所述，我们提出螺旋槽气动推力作为一种粘性真空泵，将使我们能够降低在激光扫描仪多棱镜风阻电机功率损耗。此外，我们对抽水和承载特性进行了理论和实验研究，勇两种数值计算方法对滑流进行研究，第一种方法采用的是惠普尔（1958年），Vohr及周（1965年）和Malanoski潘（1965）的窄槽理论报告。第二种方法是川端（1986年）使用边界拟合坐标系统分歧配方（DF）的方法。因此，这项工作的目标是要证实在理论和实验研究后激光扫描马达轴承有用之处。2、使用螺旋槽气动推力轴承扫描马达作为一种粘性的真空泵 图3显示了激光扫描仪采用螺旋槽气动推力轴承运转马达作为一种粘性真空泵的结构，轴的位置是固定的，转子和镜是旋转的。两个相同的螺旋槽气动推力轴承位于转子的上部和下部，在推力轴承的内圈，钻有几个小孔以致泵中的空气可以通过这个漏洞来进行抽吸。这个空间可以保持在真空中密封紧密。图4显示了螺旋槽推力轴承的几何构型，以及本文所用的符号，许多螺旋槽带有倾斜角，b形成甚至间隔于轴承圆周方向的表面上。此外，半径最短长度，lw（其中lw= R1- R2）的建立，使这两个轴承泵支持转子和镜子的重量。3、数值计算方法 两种计算方法被建议用来获取推力轴承抽水的特性数值，第一种方法用的是窄槽理论和第二种方法是用边界拟合坐标系的测向方法。此外，Burgdorfer建议（1959年一阶滑移流）在计算轴承间隙中。3.1窄槽理论在由Malanoski和潘（1965）提出的窄槽理论中，假设螺旋槽数是无限的，而宽度是微不足道的。由于在圆周方向没有压力梯度，在螺旋槽的质量流率在R-方向区域是 在内圈的推力轴承质量流率为 为了获得在压力分布的数值计算，在轴承间隙的连续性方程中的一个小元素在使用方程被进行了假设(1)和 (3) 派生方程数值的求解，假设在下面的边界条件， （4）在其中是为最终在没有压力流量在轴承间隙存在的定义。因此，最终的压力，可以通过求解方程。 （1）根据边界条件，有q=0。此外，当一个负载被施加于转子上，根据流量连续性方程就可以得到空间的压力。3.2DF法边界拟合坐标系统在窄槽理论，槽数假设是无限的，因此沟槽数目的抽水特点是用这一理论并不适用。因此，我们使用了边界拟合坐标系中的DF方法计算凹槽数方法。在边界安装质量流量坐标系，n-g坐标，如图所示5，通过转化，推导出由川端康成（1986）提出R -坐标。对于质量流量在方向 对于在方向质量流量 图5显示了清除控制体积在轴承里的质量流量连续性假设。因此，在正常的质量流率方向的常数或边界是由下列公式给出来源于连续性质量和流量3.3计算在一个密封容积随着时间的推移压力的变化转子开始转动后在密闭容积中的压力计算公式（1）假设没有轴向负荷强加给转子，并且上下轴承间隙相同，然后通过窄槽理论，质量流量在单位时间内给出率为假设在同一时刻一定体积的空气中容积的质量为G(t)，在等温条件下质量的增量表达式为：此外式中的压力P密度q 容积中在时间t时刻时的质量为m，由下式给出： 将式（11）代入式（10）中有在时间t + Dt,时刻的压力如下：在计算当中我们假设T= 0.0167 并且在T=0时刻容积的初始压力定位大气压力，其中的压力包含在（1）条件下（12）中受到的外界条件影响4、计算的结果 如上所述，窄槽理论不能用于估计螺旋槽轴承泵凹槽数量影响的特点，虽然这个理论非常简单但是对于设计气动螺旋槽是非常有用的。因此，窄槽理论是否适合螺旋槽推力轴承抽水特性预测通过在这纸上的对比获得的数据证实了DF理论。在我们的计算中使用的主要尺寸见表1在图6中用边界拟合坐标系的方法对n=8和15时的压力分布图DF理论可以用来对凹槽的形状进行考虑计算如图6因此，可以看出，在圆周方向凹槽压力分布变化的情况。通过增加数量的凹槽，凹槽之间的下跌压力变化，通过预见压力分布的方法，将获得使用窄槽理论。图7把极限压力下对使用窄槽理论与用DF理论考虑轴承间隙做了一个对比，使用df方法沟槽的数量从6改变到15，如图7中不考虑滑流效应的理论使用窄槽理论也可以参考。正如图中可以看到使用DF理论n=15的最终压力与采用窄沟理论基本相同最终压力随着沟槽数量的减少而增加。然而如图7显示沟槽数量的效应并不是很大。从这些意见中，因为螺旋槽轴承n= 15被用于实验然后使用窄槽进行理论计算和讨论。图8显示了理论压力分布在0.4个轴向压力的影响以减小内圈的承载能力。在没有轴承间隙的地方内圈的地方压力分布差异较小，这种情况下只能产生消极的负载能力。与此相反在大型轴承中这样的差异较大。此外，极限压力，聚氨酯对内圈区域不敏感。图9轴向位移和负载能力在20000转时之间的关系。传统的轴承在大气压力下操作与轴承的负荷能力相比较，轴承的负载能力比传统的那些较小但是它足以支持扫描仪转子的运作。为0.1时的承载能力相当于约负荷4N的影响，5、对比实验结果 为了验证计算结果，我们进行了一系列实验图10显示了仪器用于测量泵和拟议轴承的负荷能力实验，多边形镜子是一个重视的转子这是由三气静压轴承支撑的。多边形镜子是位于两个相同的气动螺旋槽推力轴承，镜子和推力轴承陶瓷制造的。多边形的镜子外层空间是连接到一个有六十五立方厘米量密封管容积。容积内部的压力测量采用真空压力表，推力轴承轴承间隙使用间隔由图板甲表示不同厚度来改变如图10，转子是使用直流电动机操作的。图11显示了经过旋翼开始旋转后，内部压力之间的关系。转子开始转动后五分钟后容积内部的压力迅速下降。图12显示了当转速和轴承间隙发生时变化时的极限压力正如图12中可以看到：当容积压力0.01 MPa在20000转的速度运行时提出用螺旋槽推力轴承可以使负载明显下降。图13显示了容积的极限压力负载变化对轴承的影响。在这个实验中，转子=180克的质量被用于施加在轴承上的负载。该试验台是设置在一个电磁吸盘其在水平方向具有一个倾斜角度。正如图13中可以看到最终的压力不会受到负荷的巨大影响，并且提出轴承可以支持在没有任何附加影响下转子的特点。6、结论 我们已经开发出一种螺旋槽气动推力轴承作为真空水泵轴承的运作的一种激光马达，以减少多边形的定子和风阻功率的损耗。并对轴承的特性进行了理论与实验的验证研究。因此，推导出下面的结论：1、 这种设想在hr0= 2.5um在2000转时螺旋旋槽减少了轴承密封容积的压力。2、 虽然拟议的轴承承载能力相对于大气压力轴承较小，但是它足以支持这种类型的转子质量。3、提出了可以很好地预测影响抽水性能的计算方法- 6 -TECHNICAL PAPERA method of reducing the windage power loss in a laser scannermotor using spiral-groove aerodynamic thrust bearingsfunctioning as a viscous vacuum pumpShigeka Yoshimoto Masaaki Miyatake Tomoatsu Iwasa Akiyoshi TakahashiReceived: 29 June 2006/Accepted: 2 November 2006/Published online: 1 December 2006?Springer-Verlag 2006AbstractWe propose a spiral-groove aerodynamicthrust bearing functioning as a viscous vacuum pump ina laser scanner motor to reduce the windage power lossof a polygon mirror. The proposed bearing pumps outthe air in the scanner housing using the pumping effectof the spiral-groove thrust bearing, reducing the innerpressure of the housing. The pumping performancesand the static characteristics of the spiral-groove thrustbearings were investigated numerically and experi-mentally. Two numerical calculation methods wereused to study the pumping characteristics of the spiral-groove thrust bearing. It was found that a bearing with15 spiral grooves reduced the inner pressure of thehousing to 0.01 in the bearing clearance, andthat the slip flow in the bearing clearance influencedthe pumping characteristics of the viscous vacuumpump.As mentioned above, we proposed using a spiral-groove aerodynamic thrust as a viscous vacuum pump,which would enable us to reduce the windage powerloss of the polygon mirror in a laser scanner motor, andreduce the size of the laser scanner motor. In addition,the pumping and load-carrying characteristics of ourproposed bearing were investigated theoretically andexperimentally. Two numerical calculation methodsconsidering the slip flow were used. The first methodemployed was the narrow-groove theory reported byWhipple (1958), Vohr and Chow (1965) and Malanoskiand Pan (1965), and the second method used was thedivergenceformulation(DF)methodusingtheboundary-fitted coordinate system proposed by Ka-wabata (1986). Therefore, the objective of this workwas to confirm the usefulness of our proposed bearingfor a laser scanner motor, both theoretically andexperimentally.2 The proposed scanner motor using a spiral-grooveaerodynamic thrust bearing as a viscous vacuumpumpFigure 3 shows the structure of a laser scanner motorusing spiral-groove aerodynamic thrust bearings func-tioning as a viscous vacuum pump. The shaft is fixed tothe housing, and the rotor and mirror rotate. Twoidenticalaerodynamicthrustbearingswithspiralgrooves are located at the upper and lower parts of therotor. In the inner circle of the thrust bearing, severalsmall holes are drilled so that the air inside the housingcan be pumped out through these holes. The housing issealed tightly to maintain a high vacuum in the hous-ing.Figure 4 shows the geometrical configuration of thethrust bearing with spiral grooves, along with thesymbols used in this paper. A number of spiral grooveswith groove tilt angle, b, are formed at even intervals inthe circumferential direction on the bearing surface. Inaddition,alandregion,withlength,lw(wherelw= r1 r2), is created, allowing the bearing to bothpump air and support the weight of the rotor andmirror.3 Numerical calculation methodsTwo calculation methods were used to obtain thepumping characteristics of the proposed thrust bearingnumerically. The first method used was narrow-groovetheory, and the second method was the DF methodwith a boundary-fitted coordinate system. Further-more, the first-order slip flow proposed by Burgdorfer(1959) was considered in the bearing clearance in bothcalculations.3.1 Narrow-groove theoryIn narrow-groove theory presented by Malanoski andPan (1965), it is assumed that the number of spiralgrooves 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-grooveregion in the r-direction isq ?pRTk1dpdr? k2rcosb?rdh;1wherek0 1 ? aAg aArk1AgAr a1 ? aAg? Ar2sin2bno.12lk0k2 xa1 ? ahg? hrAg? Arsinb?2k0andAg h3g1 6khg?;Ar h3r1 6khr?:2For the land region in the inner circle of the thrustbearing, the mass flow rate isq ?p12lRTh3r1 6khr?dpdrrdh:3To obtain the distribution in pressure in numericalcalculations, the equation of continuity is assumed in asmall element in the bearing clearance using Eqs. (1)and (3). The derived equation is solved numericallyassuming the following boundary conditions, atMicrosyst Technol (2007) 13:112311301125123r r2; p pa; and at r r0; p pu;4where puis defined as the ultimate pressure where noflow rate exists in the bearing clearance. Therefore, theultimate pressure, pu, can be obtained by solvingEq. (1) under the boundary condition where q = 0.Furthermore, when a load is imposed on the rotor, thecontinuity of mass flow rate is assumed betweenthe upper and lower bearing clearances, and thereforethe inner housing pressure can be obtained.3.2 DF method with boundary-fitted coordinatesystemIn narrow-groove theory, the number of grooves isassumed to be infinite and hence, the effect of thenumber of grooves on the pumping characteristics isnot applicable using this theory. Therefore, we usedthe DF method with a boundary-fitted coordinatesystem to clarify the effect of the number of grooves.The mass flow rates in the boundary-fitted coordinatesystem, n g coordinates, as shown in Fig. 5, werederived by transformation from the polar r h coor-dinates as presented by Kawabata (1986). For the massflow rate in the n-directionqnpRTffiffiffiap?Apn Bpg D?:5For the mass flow rate in the g-directionqgpRTffiffifficpBpn? Cpg E?6whereA h312l1 6kh?aJB h312l1 6kh?bJC h312l1 6kh?cJD ?rx2hrgE rx2hrna rg?2rhg?2b rnrgrhnrhgc rn?2rhn?2J rnrhg?rgrhn:Figure 5 shows a control volume in the bearingclearance where a continuity of the mass flow rate wasassumed. Therefore, the mass flow rate in the normaldirection to the constant n or g boundary is given byQnZg2g1pRT?Apn Bpg D?dgQgZn2n1pRTBpn? Cpg E?dn:7The following equation is derived from the continuityof mass flow rateRotorPump outSealed housingPolygon mirrorMotorPump outExhaust holesThrust bearing with spiral groovesJournal bearingFig. 3 A laser scanner motor using spiral-groove aerodynamicthrust bearings as a viscous vacuum pumpGrooveRidgeShaftRotorPump inPump inGroove er0r1r2hdGrooveRidgeLandThrust bearingExhaust holeshrhdPump outrhr0region1- aFig. 4 The configuration of the spiral grooves and symbols1126Microsyst Technol (2007) 13:11231130123Qn2A1 Qn1A3? Qn2A2? Qn1A4 Qg2A1 Qg1A2? Qg2A3? Qg1A4 0:83.3 Calculation of the pressure in a sealed housingwith timeThe pressure in the sealed housing was calculated usingEq. (1) after the rotor had begun to rotate. Assumingthat no axial load was imposed on the rotor, and thatthe upper and lower bearing clearances had the samevalue, then, in the narrow-groove theory, the mass flowrate in unit time is given byQt 2Z2p0qjrr2r2dh:9Next, assuming that the mass of air in the housing witha volume, Vi,was G(t) at time t, then, under isothermalconditions, G(t + Dt) is expressed asG t Dt G t ? Q t Dt:10Furthermore, using pressure, pi, and density, qi, in thehousing, the mass of air included in the housing at timet, is given byG t qit Vi pit Vi=RT*qit pit =RT:11SubstitutingEq. (11)intoEq. (10),thehousingpressure at time, t + Dt, is given as follows.pit Dt pit ? Q t RTDt=Vi:12In our calculations, Dt = 0.0167 s, and at t = 0, theinitial value of the housing pressure is assumed to beatmospheric pressure. The housing pressure obtainedby Eq. (12) is used as the boundary condition ofEq. (1) for the next time step.4 Results of the calculationsAs mentioned above, narrow-groove theory cannot beused to estimate the effect of the number of grooves onthe pumping characteristics of the spiral-groove bear-ing, although this theory is very simple, and is veryuseful for designing aerodynamic bearings with her-ringbone or spiral grooves. Therefore, the suitability ofnarrow-groove theory to predict the pumping charac-teristics of the spiral-groove thrust bearing treated inthis paper was confirmed by comparing data obtainedusing the DF method. The principal dimensions used inour calculations are shown in Table 1.Figure 6 shows the pressure distribution in thebearing clearances obtained using the DF method witha boundary-fitted coordinate system for n = 8 and 15.The DF method can be used to calculate the pressuredistribution by considering the groove shape, as shownin Fig. 6. Accordingly, it can be seen that the pressuredistribution varied between the grooves in the cir-cumferential direction. By increasing the number ofgrooves, the pressure variation between the groovesdecreased, and it can be predicted that the pressuredistribution would approach that obtained using nar-row-groove theory.Figure 7 shows a comparison of the theoreticalultimate pressure obtained using narrow-groove theoryand the DF method considering the slip flow in thebearing clearance. The number of grooves was changedfrom 6 to 15 using the DF method. In Fig. 7, a theo-retical result without considering the slip-flow effectusing narrow-groove theory is also shown for refer-ence. As can be seen in Fig. 7, the ultimate pressure forn = 15 using the DF method had almost the same valueas that using narrow-groove theory, and the ultimatepressure increased with decreasing the number ofgrooves. However, the effect of the number of grooveswas not large under the conditions shown in Fig. 7.From these observations, the theoretical calculationsdiscussed below were carried out using narrow-groovetheory because the spiral-groove bearing with n = 15was used in the experiment.A1A3A4A2 iiij jj Q2A1Q1A2Q2A3Q1A4 Q2A1Q1A3Q2A2Q1A4rqxxxxxxhhhhhhFig. 5 The n and g coordinates and mass flow continuity in acontrol volumeTable 1 Principal dimensions of the spiral-groove thrust bearingr0(mm)r1(mm)r2(mm)ab (?)hd(lm)n14.09.08.50.515.012.015Microsyst Technol (2007) 13:112311301127123Figure 8 shows the theoretical pressure distributionwith and without the inner land region for a dimen-sionless axial displacement of 0.4 to clarify the effect ofthe inner land region on the load capacity. When therewas no inner land region, the difference between thepressure distributions in the upper and lower bearingclearances was very small, and in this case even anegative load capacity was generated. In contrast, thedifference became large in the bearings with an innerland region. In addition, the ultimate pressure, Pu, wasnot sensitive to the presence of an inner land region.Figure 9showstherelationshipbetweenthedimensionless axial displacement and the dimension-less load capacity at 20,000 rpm. In Fig. 9, the loadcapacity of the proposed bearing is compared to that ofa conventional bearing that was operated at atmo-spheric pressure. The load capacity of the proposedFig. 6 Pressure distributionobtained using the DFmethod and spiral-grooveshapes010000200003000000.20.40.60.81Rotational Speed : N rpmUltimate Pressure : Pua=0.5b = 15deg.hr0=2.5 mDF Method (Slip Flow)n= 6815Narrow Groove TheoryNo Slip FlowSlip FlowFig. 7 Effect of the number of spiral grooves on the ultimatepressure in the housing0.810.511.50Dimensionless Pressure PR CoordinateLandRegionGroove Region : With Land Region lw=0.5mm : No Land Regionhr0=4m : 30000rpm :=0.4Lower BearingPressureUpper BearingPressureFig. 8 Effect of the inner land region on the pressure distribu-tion1128Microsyst Technol (2007) 13:11231130123bearing showed relatively smaller values compared tothose of the conventional bearing, but it was largeenough to support the scanner rotor. The dimension-less load capacity of 0.1 corresponds to a load of about4N in the proposed bearing.5 Comparison with the experimental resultsTo verify the calculated results, we conducted a seriesof experiments. Figure 10 shows the experimentalapparatus used to measure the pumping characteristicsand the load capacity of the proposed bearing. Thepolygon mirror was attached to one end of the rotor,which was supported by three aerostatic journal bear-ings. The polygon mirror was located between twoidenticalaerodynamicthrustbearingswithspiralgrooves. The mirror and the thrust bearings were madeof ceramic. The outer space of the polygon mirror wasconnected by a tube to a tightly sealed housing with avolume of 65 cm3. The pressure inside the housing wasmeasured using a vacuum pressure gauge. The bearingclearance of the thrust bearing could be changed usingspacers with different thicknesses, denoted by Plate Ain Fig. 10. The rotor was operated using a DC motor.Figure 11 shows the relationship between the innerpressure of the housing and the elapsed time after therotor had begun rotating. The inner pressure decreasedrapidly immediately after the rotor had begun rotating,and leveled out within a period of 5 min.0. 20. 40. 60.80.20.40.60.80Dimensionless Axial Displacement Dimensionless load-carrying capacity W20000rpmPresent Conv.hr0=3.0m :hr0=4.0m :hr0=5.0m :Fig. 9 The relationship between the axial displacement and theload capacityTubeDigital TachometerPolygon MirrorPlate AMotorHousingAerostatic Journal BearingsShaftSpiral-Grooved Thrust BearingPressure GaugeVolumeAir outAir outAir inFig. 10 The experimental apparatus10200.20.40.60.810Cal. Exp. : hr0=2.5 m : hr0=3.5 m : hr0=4.5 mElapsed time : t minDimensionless Pressure in the Chamber Pilw=0.5mm : 20000rpmPuFig. 11 Variation of the housing pressure with elapsed time1000020000300000.20.40.60.810Rotational speed N rpmUltimae Pressure PuExp. Cal. : hr0=2.5m : hr0=3.5m : hr0=4.5mFig. 12 Effect of the rotational speed and bearing clearance onthe ultimate pressureMicrosyst Technol (2007) 13:112311301129123Figure 12 shows the ultimate pressure when therotational speed and the bearing clearance werechanged. As can be seen in Fig. 12, the proposed thrustbearing with spiral grooves decreased the inner pres-sure of the housing to 0.01 MPa at hr0= 2.5 lm at aspeed of 20,000 rpm.Figure 13 shows the variation in the ultimate pres-sure in the housing when a load was imposed on thebearings. In this experiment, a rotor mass = 180 g wasused to impose a load on the bearings. The test rig wasset on a magnetic chuck, which could be tilted from thehorizontal to a vertical angle. As can be seen in Fig. 13,the ultimate pressure was not greatly influenced by theimposed load, and the proposed bearing could supportthe rotor without any deterioration in the pumpingcharacteristics.6 ConclusionsWe have developed a spiral-groove aerodynamic thrustbearing functioning as a vacuum pump to reduce thewindage power loss of the polygon mirror and the sizeof a laser scanner motor. The pumping characteristicsof the proposed bearing were investigated theoreticallyand experimentally. As a result, the following conclu-sions were derived:1.The proposed bearing with spiral grooves reducedthe pressure in the sealed housing to 0.01 MPa athr0= 2.5 lm at 20,000 rpm.2.Though the load capacity of the proposed bearingwas relatively small compared with that of abearing operating at atmospheric pressure, it waslarge enough to support the rotor mass in
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