機械專業(yè)畢業(yè)論文外文翻譯--實驗研究激光加工表面微觀造型平行的推力軸承

1、<p>　　附錄一 英文科技文獻翻譯</p><p><b>　　英文原文：</b></p><p>　　Experimental investigation of laser surface textured parallel thrust bearings</p><p>　　Performance enhancements b

2、y laser surface texturing (LST) of parallel-thrust bearings is experimentally investigated. Test</p><p>　　results are compared with a theoretical model and good correlation is found over the relevant operati

3、ng conditions. A compari-</p><p>　　son of the performance of unidirectional and bi-directional partial-LST bearings with that of a baseline, untextured bearing is</p><p>　　presented showing the

4、bene?ts of LST in terms of increased clearance and reduced friction.</p><p>　　KEY WORDS: ?uid ?lm bearings, slider bearings, surface texturing</p><p>　　1. Introduction</p><p>　　The

5、classical theory of hydrodynamic lubrication yields linear (Couette) velocity distribution with zero pressure gradients between smooth parallel surfaces under steady-state sliding. This results in an unstable hydrodynami

6、c ?lm that would collapse under any external force acting normal to the surfaces. However, experience shows that stable lubricating ?lms can develop between parallel sliding surfaces, generally</p><p>　　beca

7、use of some mechanism that relaxes one or more of the assumptions of the classical theory.</p><p>　　A stable ?uid ?lm with su?cient load-carrying capacity in parallel sliding surfaces can be obtained, for ex

8、ample, with macro or micro surface structure of di?erent types. These include waviness [1] and protruding microasperities [2–4]. A good literature review on the subject can be found in Ref. [5]. More recently, laser surf

9、ace texturing (LST) [6–8], as well as inlet roughening by longitudinal or transverse grooves [9] were suggested to provide load capacity in parallel sliding. The inlet rough</p><p>　　direction and in this re

10、spect it is identical to the par- tial-LST concept described in ref. [10] for generating hydrostatic e?ect in high-pressure mechanical seals.</p><p>　　Very recently Wang et al. [11] demonstrated experimental

11、ly a doubling of the load-carrying capacity for the surface- texture design by reactive ion etching of SiC parallel-thrust bearings sliding in water. These simple parallel thrust bearings are usually found in seal-less p

12、umps where the pumped ?uid is used as the lubricant for the bearings. Due to the parallel sliding their performance is poorer than more sophisticated tapered or stepped bearings. Brizmer et al. [12] demon-strated the po

13、tent</p><p>　　was analyzed. The optimum parameters of the dimples were found in order to obtain maximum load-carrying capacity. A micro-dimple ‘‘collective e?ect’’ was identi-</p><p>　　?ed that

14、is capable of generating substantial load-carrying capacity, approaching that of optimumconventional thrust bearings. The purpose of the present paper is to investigate experimentally the validity of the model described

15、in Ref. [12] by testing practical thrust bearings and comparing the performance of LST bearings with that of the theoretical predictions and with the performance of standard non-textured</p><p><b>　　be

16、arings</b></p><p>　　2. Background</p><p>　　A cross section of the basic model that was analyzed in Ref. [12] is shown in figure </p><p>　　1. A slider having a width B is parti

17、ally textured over a portion Bp =αB of its width. The textured surface consists of multiple dimples with a diameter，depthand area density Sp. As a result of the hydrodynamic pressure generated by the dimples the sliding

18、surfaces will be separated by a clearancedepending on the sliding velocity U, the ?uid viscosity l and the external loadIt was found in Ref. [12] that an optimum ratio exists for the parameter that provides maximum dimen

19、sionless load-carrying </p><p>　　the bearing length, and this optimum value is hp=1.25. It was further found in Ref. [12] that an optimum value exists for the textured portion a depending onthe bearing aspec

20、t ratio L/B. This behavior is shown in ?gure 2 for a bearing with L/B = 0.75 at various values of the area density Sp. As can be seen in the range of Sp values from 0.18 to 0.72 the optimum a value varies from 0.7 to 0.5

21、5, respectively. It can also be seen from ?gure 2 that for a < 0.85 no optimum value exists for Sp and the</p><p>　　3. Experimental</p><p>　　The tested bearings consist of sintered SiC disks

22、10 mm thick, having 85 mm outer diameter and 40 mm inner diameter. Each bearing (see ?gure 3) comprises a ?at rotor (a) and a six-pad stator (b). The bearings were provided with an original surface ?nish</p><p

23、>　　by lapping to a roughness average Ra= 0.03 lm. Each pad has an aspect ratio of 0.75 when its width is measured along the mean diameter of the stator. The photographs of two partial-LST stators are shown in ?gure 4

24、where the textured areas appear as brighter matt surfaces. The ?rst stator indicated (a) is a unidirectional bearing with the partial-LST adjacent to the leading edge of each pad, similar to the model shown in ?gure 1. T

25、he second stator (b) is a bi-directional version of a partial-LST </p><p>　　electrical motor turns a spindle to which an upper holder of the rotor is attached. A second lower holder of the stator is ?xed to

26、a housing, which rests on a journal bearing and an axial loading mechanism that can freely move in the axial direction</p><p>　　. An arm that presses against a load cell and thereby permits friction torque m

27、easurements prevents the free rotation of this housing. Axial loading is provided by means of dead weights on a lever and is measured with a second load cell. A proximity probe that is attached to the lower holder of the

28、 stator allows on-line measurements of the clearance change between rotor and stator as the hydrodynamic e?ects cause axial movement of the housing to which the stator holder is ?xed. Tap water is supp</p><p&g

29、t;　　the outer diameter of the bearing allows monitoring of the water temperature as the water exit the bearing. A PC is used to collect and process data on-line. Hence,the instantaneous clearance, friction coe?cient, bea

30、ring speed and exit water temperature can be monitored constantly.</p><p>　　The test protocol includes identifying a reference “zero” point for the clearance measurements by ?rst loading and then unloading a

31、 stationary bearing over the full load range. Then the lowest axial load is applied, the water supply valve is opened and the motor turned on. Axial loading is increased by steps of 40 N and each load step is maintained

32、for 5 min following the stabilization of the friction coe?cient at</p><p>　　a steady-state value. The bearing speed and water temperature are monitored throughout the test for any irregularities. The test en

33、ds when a maximum axial load of 460 N is reached or if the friction coe?cient exceeds a value of 0.35. At the end of the last load step the motor and water supply are turned o? and the reference for the clearance measure

34、ments is rechecked. Tests are performed at two speeds of 1500</p><p>　　and 3000 rpm corresponding to average sliding velocities of 4.9 and 9.8 m/s, respectively and each test is repeated at least three times

35、.</p><p>　　4. Results and discussion</p><p>　　As a ?rst step the validity of the theoretical model in Ref. [12] was examined by comparing the theoretical and experimental results of bearing clea

36、rance versus bearing load for a unidirectional partial-LST bearing. The results are shown in ?gure 6 for the two speeds of 1500 and 3000 rpm where the solid and dashed lines correspond to the model and experiment, respec

37、tively. As can be seen, the agreement between the model and the experiment is good, with di?erences of less than 10%, as long as the</p><p>　　It should be noted here that the ?rst attempts to test the baseli

38、ne untextured bearing with the original surface ?nish of Ra= 0.03 lm on both the stator and rotor failed due to extremely high friction even at the lower loads. On the other hand the partial-LST bearing ran smoothly thro

39、ughout the load range. It was found that the post-LST lapping to completely remove about 2 lm height bulges, which are formed during texturing around the rims of the dimples, resulted in a slightly rougher surface w</

40、p><p>　　ness of the partial-LST stator and all subsequent tests were performed with the same Ra value of 0.04 lm for all the tested stators. The rotor surface roughness</p><p>　　remained, the origi

41、nal one namely, 0.03 lm. Figure 7 presents the experimental results for the clearance as a function of the load for a partial-LST unidirectional bearing (see stator in ?gure 4(a)) and a baseline untextured bearing. The c

42、omparison is made at the two speeds of 1500 and 3000 rpm. The area density of the dimples in the partial-LST bearing is Sp= 0.6 and the textured portion is a ¼ 0:734. The load range extends from 160 to 460 N. The up

43、per load was determined by the test-rig limita</p><p>　　and 2.2 lm for the LST and untextured bearings, respectively. As can be seen from ?gure 7 this ratio of about 3 in favor of the partial-LST bearing is

44、maintained over the entire load range.</p><p>　　Figure 8 presents the results for the bi-directionalbearing (see stator in ?gure 4(b)). In this case the LST parameters are Sp ¼ 0:614 and a ¼ 0:633.

45、 The clearances of the bi-directional partial-LST bearing are lower compared to these of the unidirectional bearing at the same load. At 460 N load the clearance for the 1500 rpm is 4.1 lm and for the 3000 rpm it is 6 lm

46、. These values represent a reduction of clearance between</p><p>　　33 and 10% compared to the unidirectional case. However, as can be seen from ?gure 8 the performance of the partial-LST bi-directional beari

47、ng is still substantially better than that of the untextured bearing.</p><p>　　The friction coe?cient of partial-LST unidirectional and bi-directional bearings was compared with that of the untextured bearin

48、g in ?gures 9 and 10 for the two speeds of 1500 and 3000 rpm, respectively. As can be seen the friction coe?cient of the two partial-LST bearings is very similar with slightly lower values in the case of the more e?cient

49、 unidirectional bearing. The friction coe?cient of the untextured bearing is</p><p>　　much larger compared to that of the LST bearings. At 1500 rpm (?gure 9) and the highest load of 460 N the friction coe?ci

50、ent of the untextured bearing is about 0.025 compared to about 0.01 for the LST bearings.</p><p>　　At the lowest load of 160 N the values are about 0.06 for the untextured bearing and around 0.02 for the LST

51、 bearings. Hence, the friction values of the untextured bearing are between 2.5 and 3 times higher than the corresponding values for the partial-LST bearings over the entire load range. Similar results were obtained at t

52、he velocity of 3000 rpm (?gure 10) but the level of the friction coe?cients is somewhat higher</p><p>　　due to the higher speed. The much higher friction of the untextured bearing is due to the much smaller

53、clearances of this bearing (see ?gures 7 and 8) that result in higher viscous shear.</p><p>　　5. Conclusion</p><p>　　The idea of partial-LST to enhance performance of the parallel thrust bearing

54、 was evaluated experimentally. Good correlation was found with a theoretical model as</p><p>　　long as the basic assumption of laminar ?ow in the ?uid ?lm is valid. At low loads with relatively large clearan

55、ces, where turbulence may occur, the experimental</p><p>　　clearance is larger than the prediction of the model.The performance of both unidirectional and bidirectional partial-LST bearings in terms of clear

56、ance</p><p>　　and friction coe?cient was compared with that of a baseline untextured bearing over a load range in which the theoretical model is valid. A dramatic increase, of</p><p>　　about thr

57、ee times, in the clearance of the partial-LST bearings compared to that of the untextured bearing was obtained over the entire load range. Consequently the friction coe?cient of the partial-LST bearings is much lower, re

58、presenting more than 50% reduction in friction compared to the untextured bearing.</p><p>　　The larger clearance and lower friction make the partial-LST simple parallel thrust bearing concept much more relia

59、ble and e?cient especially in seal-less pumps and similar applications where the process ?uid, which is often a poor lubricant, is the only available lubricant for the bearings.</p><p>　　Acknowledgments</

60、p><p>　　The authors would like to thank Mr. J. Boylan of Morgan AM&T for providing the bearing specimens and Mr. N. Barazani of Surface Technologies Ltd. For providing the laser surface texturing. </p>

61、;<p>　　實驗研究激光加工表面微觀造型平行的推力軸承</p><p>　　實驗是研究激光處理的表面微觀造型平行的推力軸承增強的某些性能。測試結(jié)果與理論模型進行了比較,，發(fā)現(xiàn)在有關(guān)的運行條件之外有著別的關(guān)系。突出表現(xiàn)在，單向和雙向定向部分反演軸承與一個基線的關(guān)系，激光表面微觀造型與無微觀造型軸承的比較顯示好處在于，增加了清理和減少摩擦。</p><p>　　關(guān)鍵詞：油膜軸承

62、，滑塊，軸承，表面微觀造型</p><p><b>　　緒論</b></p><p>　　經(jīng)典理論的流體動力潤滑產(chǎn)生線性（ couette ）的速度分布與零壓力梯度之間的順利進行平行表面下的穩(wěn)定狀態(tài)滑動。這個結(jié)果在不穩(wěn)定的潤滑膜在任何外部力在表面起作用的情況下會破裂。不過，經(jīng)驗表明，穩(wěn)定的潤滑膜可以擴大他們之間的平行滑動面，一般由于某些機制，放寬一種或一種以上的對經(jīng)典

63、理論的假設(shè)。</p><p>　　在平行滑動面可以得到一個穩(wěn)定的，有足夠的承載能力的油膜，例如，宏觀或微觀表面結(jié)構(gòu)就是不同類型。這些措施包括波紋形[ 1 ]和凸起微粗糙面[ 2-4 ]。一個好的工藝系統(tǒng)就是一個標(biāo)準(zhǔn) [ 5 ] 。最近，激光表面紋理 [ 6-8 ] ，就是開口粗糙的縱向或橫向的凹槽[ 9 ]在平行滑動提供承載能力。開口粗糙度的概念既[ 9 ]是基于有效地清除，減少在滑動方向和在這方面是相同的部分激

64、光表面微造型概念所描述的標(biāo)準(zhǔn)。[ 10 ]產(chǎn)生靜壓力對高壓力的機械密封影響。最近，王等人。 [ 11 ]實驗表明，增加一倍的承載能力為表面紋理設(shè)計的反應(yīng)離子刻蝕碳化硅平行推力軸承滑動在水中。這些簡單的平行推力軸承，通常發(fā)現(xiàn)，在密封泵少的地方抽液是用來作為潤滑劑的軸承。由于平行滑動他們的表現(xiàn)較差，比更先進的錐形或加強軸承。brizmer等人。 [ 12 ]表現(xiàn)出的潛力，激光表面紋理在的形式，定期微量波紋提供承載能力與平行推力軸承。模型的紋

65、理平行滑塊是發(fā)達(dá)國家和作用的表面紋理對承載能力進行了分析。最佳參數(shù)的微波被發(fā)現(xiàn)，以取得最大的承載能力。微蜂窩集體效應(yīng)被鑒定是能產(chǎn)生可觀的承載能力，接近的最佳的傳統(tǒng)推力軸承。該本文</p><p><b>　　第二章 背景</b></p><p>　　基本模型的橫截面用標(biāo)準(zhǔn)分析了[ 12 ]是表現(xiàn)在圖1?；瑝K有一個寬度B是部分微觀造型BP = αB的寬度。該紋理的表面

66、組成眾多波紋同一的直徑為 深度為分布密度為自身屬性。人們發(fā)現(xiàn)，有著微觀表面造型的滑動面的油壓被分開是與滑動速度U、液體粘度1和外部負(fù)載W有關(guān)[ 12 ]認(rèn)為，有一個最佳的比例參數(shù)存在能使微觀表面造型提供最大的無量綱負(fù)載。其中L是軸承的長度，且最浩的動力是HP=1.25.</p><p>　　這是進一步發(fā)現(xiàn)， [ 12 ]認(rèn)為，部分的表面微觀造型存在一個最佳值為軸承長寬比L/B這種行為是如圖2所示為軸承 b = 0

67、.75在不同的價值觀該地區(qū)的密度藻可以看出，在從0.18至0.72范圍內(nèi)發(fā)現(xiàn)SP值的最佳值不同，分別從0.7至0.55 。它也可以從圖2 ，對于一個0.85<密度是沒有最優(yōu)值的SP存在且最高負(fù)荷瓦特與SP同步增加，因此，最大的面積密度，可以得到幾乎與激光毛化是理想的。這亦是有趣地注意到，從圖2，我們看到用軟件仿真的部分表面微觀造型的優(yōu)勢。舉例說明，在SP=0.5比例α=0.6時是α=1時的三倍的的承載能力。</p>

68、<p>　　第三章 實驗 </p><p>　　測試軸承組成燒結(jié)碳化硅磁盤10毫米厚，有八十五毫米外徑和40毫米內(nèi)徑。每個軸承（見圖3 ）組成一個單位，轉(zhuǎn)子（ a ）和6墊定子（ b ）款。軸承提供了一個原始的表面光潔度由研磨到平均粗糙度在Ra = 0.03的LM 。每個墊有一個長寬比0.75時，其寬度是衡量沿線平均直徑定子。照片2部分第1定子是如圖4所示的地方紋理地區(qū)出現(xiàn)更加美好的亞光表面。

69、第一定子表示， （ a ）是單向軸承與局部反演毗鄰的領(lǐng)先地位，每個墊，類似的模型如圖1所示。第二定子（二）是一個雙向定向版本的部分反演軸承有兩個平等的紋理部分1/2，對每一項墊結(jié)束。該激光毛化參數(shù)以下;壓痕深度，壓痕直徑和壓痕面積密度sp = 0.6 0.03 。這些壓痕的尺寸，獲得了與4脈沖30的NS的時間長短和4兆焦耳每使用1 5千赫的脈動Nd ： YAG激光。該紋理部分單向軸承是一個= 0.73和該雙向定向軸承是一個= 0.63

70、?？梢钥闯?，從圖2這兩種價值觀應(yīng)產(chǎn)生承載能力不同，接近最高的理論value.the試驗臺是顯示schematically在圖5 。電機輪流主軸，以其中一上持有轉(zhuǎn)子重視。第二個較低的持有人的定子是固定的房屋，在于</p><p>　　一個單臂反應(yīng)壓力與負(fù)載單元相互作用，從而許可證的摩擦力矩測量阻止自由旋轉(zhuǎn)這個機架。軸向載荷是所提供的手段，對絕對的權(quán)重杠桿作用，是衡量一個第二負(fù)荷單元。感應(yīng)探頭是附加到較低的持有人的定

71、子，讓上線的測量清拆變化之間的轉(zhuǎn)子和定子由于水動力影響的原因軸向運動的房屋，其中定子持有人，這是一個固定的。自來水供應(yīng)的重心從一個大罐的中心軸承和滲漏從軸承是收集和重新分發(fā)。 1熱電偶毗鄰<br>外徑軸承允許監(jiān)測水溫，作為水出口軸承。電腦是用來收集和處理數(shù)據(jù)上線。因此，瞬時關(guān)，摩擦系數(shù)，軸承的速度和開槽的溫度可不斷監(jiān)測。</p><p>　　測試草案包括確定一個參考“零”點為清除測量第一有負(fù)載和無負(fù)載

72、，然后固定軸承超過滿負(fù)荷的范圍。然后最低的軸向載荷應(yīng)用，供水閥打開及汽車開啟。軸向負(fù)荷增加的步驟40 N和每個負(fù)載的步驟是維持5分鐘之后，穩(wěn)定的摩擦系數(shù)在一穩(wěn)定狀態(tài)的價值。軸承的速度和水溫監(jiān)測整個測試的任何違規(guī)行為。試驗結(jié)束時，最大軸向負(fù)荷460 N是達(dá)到或如果摩擦系數(shù)超過了價值0.35 。在年底的最后一步負(fù)荷電機及食水供應(yīng)關(guān)掉，并參考有關(guān)清拆測量是復(fù)查。測試是在兩種速度的1500 和3000 RPM的相應(yīng)的平均滑動速度4.9和9.8米

73、/秒，分別和每個測試重復(fù)至少3次</p><p>　　第四章 成果與討論</p><p>　　作為第一步的有效性的理論模型。 [ 12 ]研究并比較，理論和實驗結(jié)果的軸承間隙銀兩軸承載荷為單向局部反演軸承。結(jié)果表明，在圖6為兩種速度的1500和3000 rpm的情況下固體和虛線對應(yīng)到模型和實驗，分別?？梢钥闯?，雙方間的協(xié)議模型和實驗是好的，與不同的不到10 ％ ，只要負(fù)荷是150以上的1

74、2月31日在較低載荷測量的實驗清拆要遠(yuǎn)遠(yuǎn)大于模型預(yù)測，尤其是在較高的速度， 3000 rpm的情況下，在120 n實測關(guān)是20的LM ，這是約60 ％ ，高于預(yù)測值。結(jié)果表明，該組合，如此龐大的間隙和相對低粘度的水可能會導(dǎo)致湍流流體膜。因此，假設(shè)油膜上，解決這一雷諾方程的標(biāo)準(zhǔn)形式。 [ 12 ]是基于可能違反決策模型無效特別是在較高的速度和最低的負(fù)荷。 [ 12 ]這是決定進一步限制比較負(fù)荷以上150 N</p><

75、p>　　它這里應(yīng)該指出，第一，企圖測試基線無微觀造型軸承與原來的表面光潔度的RA = 0.03的LM上都定子和轉(zhuǎn)子失敗，由于極高的摩擦，甚至在較低的負(fù)荷。在另一方面部分-第1軸承，整個負(fù)荷范圍順利。結(jié)果發(fā)現(xiàn)，后反演研磨完全移除約2的LM高度凸出部分 ，這是中形成的紋理周圍的輪輞的波紋 ，導(dǎo)致在一個稍微粗糙的表面粗糙度= 0.04的LM 。因此，基線與無造型的定子重疊，以同一粗糙性的部分-第1定子和其后所有測試的表現(xiàn)與定子同在Ra值

76、為0.04的LM的所有測試。轉(zhuǎn)子表面粗糙度仍然存在，原因，即0.03的LM 。圖7給出了實驗結(jié)果為清除作為一個功能負(fù)荷為局部反演單向軸承（見定子在圖4 （ a ） ）和基線無微觀造型軸承。比較是在兩種速度的1500和3000 RPM的。面密度的波紋在部分-第1軸承是sp = 0.6和紋理部分是一個6.3 0:734 。該負(fù)荷范圍擴大，從160至460 12月31日上負(fù)載檢測試驗臺的限制，不容許較高的負(fù)荷。很顯然，從圖7部分-第1軸承運轉(zhuǎn)

77、大幅清拆比無微觀造型軸承。在最高負(fù)荷460 N和速度1500 RPM的部分-第1軸承已清拆6的LM ，而無微觀造型軸承間隙是只有1.7</p><p>　　圖8給出的結(jié)果為雙向軸承（見定子在圖4 （ b ）款） 。在這種情況下，反演參數(shù)sp=6.3 α=0.614和0.633 6.3 。清拆的雙向定向部分反演軸承相比，降低這些的單向軸承在同一負(fù)荷。在460 n負(fù)載清拆為1500 rpm的是4.1 LM和為300

78、0 rpm的，這是6月的LM 。這些價值觀所代表的減少之間的關(guān)<br> 33和10 ％相比，單向的情況。不過，可以看出，從圖8的表現(xiàn)，部分-第1雙向定向軸承仍是大大優(yōu)于該無微觀造型軸承。</p><p>　　圖10為兩種速度分別是1500和3000 rpm?？梢钥闯觯Σ料禂?shù)的兩個部分反演軸承是非常類似的與略低的價值觀，在部件較有高效率的單向軸承。無微觀造型的的摩擦系數(shù)大得多比他們大的多，即第1軸承。

79、在1500 RPM的（圖9 ）和最高負(fù)荷460 n摩擦系數(shù)的untextured軸承是約0.025相比，約為第1軸承0.01。</p><p>　　在最低負(fù)荷160 n值約0.06為無微觀造型軸承的為第1軸承的0.02左右。因此，無微觀造型軸承摩擦值，高于相應(yīng)值為局部反演軸承在整個負(fù)荷范圍的2.5和3倍。，在速度上獲得了類似的結(jié)果， 3000每分鐘轉(zhuǎn)速（圖10 ） ，但水平的摩擦系數(shù)是有點高，由于較高的速度。無微

80、觀造型軸承的摩擦高得多，是因為無小槽清理磨砂（見圖7和圖8），導(dǎo)致較高的粘性剪切。</p><p>　　第五章 結(jié)束語 </p><p>　　局部觀點反映出，實驗以高度評價平行推力軸承的性能。發(fā)現(xiàn)與理論模型良好的相關(guān)性,在油膜是有效的情況下,作了大膽的基本層流假設(shè)。在低負(fù)荷與比較大的清除行動，其中可能會發(fā)生不穩(wěn)定，實驗預(yù)測的模型都表現(xiàn)了單向和雙向局部軸承與一個基線無微觀造型軸承比較,

81、其中摩擦系數(shù)急劇增加3倍左右，反負(fù)荷超過范圍,映出理論模型是有效的。 在清除部分-第1軸承相比，認(rèn)為該無微觀造型軸承，獲得了在整個負(fù)荷范圍。因此，摩擦系數(shù)的部分-第1軸承相比是要低得多，表明，減少了50 ％以上摩擦，該無微觀造型軸承。 <br>規(guī)模較大的清理和低摩擦，使局部反映簡單的平行推力軸承的概念更為可靠和高效率的，特別是在密封校水泵和類似的應(yīng)用過程中流體，這常常是一個缺少潤滑劑下，是唯一可用的潤滑劑軸承。</p&