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1、<p>　　附錄一 英文科技文獻(xiàn)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>　　Bearings fail for a number of reasons，but the most common are misapplication，contamination，improper lubricant，shi

54、pping or handling damage，and misalignment. The problem is often not difficult to diagnose because a failed bearing usually leaves telltale signs about what went wrong．</p><p>　　However，while a postmortem yie

55、lds good information，it is better to avoid the process altogether by specifying the bearing correctly in The first place．To do this，it is useful to review the manufacturers sizing guidelines and operating characteristics

56、 for the selected bearing.</p><p>　　Equally critical is a study of requirements for noise, torque, and runout, as well as possible exposure to contaminants, hostile liquids, and temperature extremes. This ca

57、n provide further clues as to whether a bearing is right for a job.</p><p>　　1 Why bearings fail</p><p>　　About 40% of ball bearing failures are caused by contamination from dust, dirt, shavings

58、, and corrosion. Contamination also causes torque and noise problems, and is often the result of improper handling or the application environment．Fortunately, a bearing failure caused by environment or handling contamina

59、tion is preventable，and a simple visual examination can easily identify the cause．</p><p>　　Conducting a postmortem il1ustrates what to look for on a failed or failing bearing．Then，understanding the mechanis

60、m behind the failure, such as brinelling or fatigue, helps eliminate the source of the problem.</p><p>　　Brinelling is one type of bearing failure easily avoided by proper handing and assembly. It is charact

61、erized by indentations in the bearing raceway caused by shock loading－such as when a bearing is dropped-or incorrect assembly. Brinelling usually occurs when loads exceed the material yield point(350,000 psi in SAE 52100

62、 chrome steel)．It may also be caused by improper assembly, Which places a load across the races．Raceway dents also produce noise，vibration，and increased torque.</p><p>　　A similar defect is a pattern of elli

63、ptical dents caused by balls vibrating between raceways while the bearing is not turning．This problem is called false brinelling. It occurs on equipment in transit or that vibrates when not in operation. In addition, deb

64、ris created by false brinelling acts like an abrasive, further contaminating the bearing. Unlike brinelling, false binelling is often indicated by a reddish color from fretting corrosion in the lubricant.</p><

65、p>　　False brinelling is prevented by eliminating vibration sources and keeping the bearing well lubricated. Isolation pads on the equipment or a separate foundation may be required to reduce environmental vibration. A

66、lso a light preload on the bearing helps keep the balls and raceway in tight contact. Preloading also helps prevent false brinelling during transit.</p><p>　　Seizures can be caused by a lack of internal clea

67、rance, improper lubrication, or excessive loading. Before seizing, excessive, friction and heat softens the bearing steel. Overheated bearings often change color，usually to blue-black or straw colored．Friction also cause

68、s stress in the retainer，which can break and hasten bearing failure．</p><p>　　Premature material fatigue is caused by a high load or excessive preload．When these conditions are unavoidable，bearing life shoul

69、d be carefully calculated so that a maintenance scheme can be worked out．</p><p>　　Another solution for fighting premature fatigue is changing material．When standard bearing materials，such as 440C or SAE 521

70、00，do not guarantee sufficient life，specialty materials can be recommended. In addition，when the problem is traced back to excessive loading，a higher capacity bearing or different configuration may be used．</p>&l

71、t;p>　　Creep is less common than premature fatigue．In bearings．it is caused by excessive clearance between bore and shaft that allows the bore to rotate on the shaft．Creep can be expensive because it causes damage to o

72、ther components in addition to the bearing．</p><p>　　0ther more likely creep indicators are scratches，scuff marks，or discoloration to shaft and bore．To prevent creep damage，the bearing housing and shaft fitt

73、ings should be visually checked．</p><p>　　Misalignment is related to creep in that it is mounting related．If races are misaligned or cocked．The balls track in a noncircumferencial path．The problem is incorre

74、ct mounting or tolerancing，or insufficient squareness of the bearing mounting site．Misalignment of more than 1/4·can cause an early failure．</p><p>　　Contaminated lubricant is often more difficult to de

75、tect than misalignment or creep．Contamination shows as premature wear．Solid contaminants become an abrasive in the lubricant．In addition。insufficient lubrication between ball and retainer wears and weakens the retainer．I

76、n this situation，lubrication is critical if the retainer is a fully machined type．Ribbon or crown retainers，in contrast，allow lubricants to more easily reach all surfaces． </p><p>　　Rust is a form of moistur

77、e contamination and often indicates the wrong material for the application．If the material checks out for the job，the easiest way to prevent rust is to keep bearings in their packaging，until just before installation．<

78、/p><p>　　2 Avoiding failures</p><p>　　The best way to handle bearing failures is to avoid them．This can be done in the selection process by recognizing critical performance characteristics．These in

79、clude noise，starting and running torque，stiffness，nonrepetitive runout，and radial and axial play．In some applications, these items are so critical that specifying an ABEC level alone is not sufficient．</p><p&g

80、t;　　Torque requirements are determined by the lubricant，retainer，raceway quality(roundness cross curvature and surface finish)，and whether seals or shields are used．Lubricant viscosity must be selected carefully because

81、inappropriate lubricant，especially in miniature bearings，causes excessive torque．Also，different lubricants have varying noise characteristics that should be matched to the application. For example，greases produce more no

82、ise than oil．</p><p>　　Nonrepetitive runout(NRR)occurs during rotation as a random eccentricity between the inner and outer races，much like a cam action．NRR can be caused by retainer tolerance or eccentricit

83、ies of the raceways and balls．Unlike repetitive runout, no compensation can be made for NRR.</p><p>　　NRR is reflected in the cost of the bearing．It is common in the industry to provide different bearing typ

84、es and grades for specific applications．For example，a bearing with an NRR of less than 0.3um is used when minimal runout is needed，such as in disk—drive spindle motors．Similarly，machine—tool spindles tolerate only minima

85、l deflections to maintain precision cuts．Consequently, bearings are manufactured with low NRR just for machine-tool applications．</p><p>　　Contamination is unavoidable in many industrial products，and shields

86、 and seals are commonly used to protect bearings from dust and dirt．However，a perfect bearing seal is not possible because of the movement between inner and outer races．Consequently，lubrication migration and contaminatio

87、n are always problems．</p><p>　　Once a bearing is contaminated, its lubricant deteriorates and operation becomes noisier．If it overheats，the bearing can seize．At the very least，contamination causes wear as i

88、t works between balls and the raceway，becoming imbedded in the races and acting as an abrasive between metal surfaces．Fending off dirt with seals and shields illustrates some methods for controlling contamination．</p&

89、gt;<p>　　Noise is as an indicator of bearing quality．Various noise grades have been developed to classify bearing performance capabilities．</p><p>　　Noise analysis is done with an Anderonmeter, which

90、is used for quality control in bearing production and also when failed bearings are returned for analysis. A transducer is attached to the outer ring and the inner race is turned at 1,800rpm on an air spindle. Noise is m

91、easured in andirons, which represent ball displacement in μm/rad.</p><p>　　With experience, inspectors can identify the smallest flaw from their sound. Dust, for example, makes an irregular crackling. Ball s

92、cratches make a consistent popping and are the most difficult to identify. Inner-race damage is normally a constant high-pitched noise, while a damaged outer race makes an intermittent sound as it rotates.</p><

93、;p>　　Bearing defects are further identified by their frequencies. Generally, defects are separated into low, medium, and high wavelengths. Defects are also referenced to the number of irregularities per revolution.<

94、;/p><p>　　Low-band noise is the effect of long-wavelength irregularities that occur about 1.6 to 10 times per revolution. These are caused by a variety of inconsistencies, such as pockets in the race. Detectabl

95、e pockets are manufacturing flaws and result when the race is mounted too tightly in multiplejaw chucks.</p><p>　　Medium-hand noise is characterized by irregularities that occur 10 to 60 times per revolution

96、. It is caused by vibration in the grinding operation that produces balls and raceways. High-hand irregularities occur at 60 to 300 times per revolution and indicate closely spaced chatter marks or widely spaced, rough i

97、rregularities.</p><p>　　Classifying bearings by their noise characteristics allows users to specify a noise grade in addition to the ABEC standards used by most manufacturers. ABEC defines physical tolerance

98、s such as bore, outer diameter, and runout. As the ABEC class number increase (from 3 to 9), tolerances are tightened. ABEC class, however, does not specify other bearing characteristics such as raceway quality, finish,