ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
races are then heat treat hardened, finish ground and polished to produce bearing quality races as shown  in Figure 5.   Figure 5.  60NiTi Bearing Races. In order to estimate the static load capacity (permanent dent resistance) of a 60NiTi hybrid bearing one  first needs to measure the load limit of a single ball-race contact. When one ball contacts a race, the load  limit is reached when the Hertz contact stress exceeds that necessary to create a permanent dent whose depth is approximately 0.00005 times that of the ball diameter. This dent depth magnitude corresponds to a damage threshold which when exceeded leads to rough bearing operation and potentially reduced fatigue life (Ref. 2). The Hertz equations are used to calculate the contact stress. The stress is a function of the load, the geometries and material properties of the ball and race. The bearing geometry evaluated  adheres to conventional design norms. The ball has a radius of 4.37 mm and the race curvatures are nominally 4.55 and 27.4 mm in the cross race and ball path (circumferential) directions, respectively. This gives the bearing ball-race conformity a value of 52 percent, a value very typical for ball bearings (Ref. 10). Since the inner race-ball contact represents the most severe stress location in a ball bearing, it  is selected in the present study as the location of interest. A special fixture was designed and fabricated to firmly hold the inner race and present it to a single  ceramic ball for loading experiments. Figure 6 shows the fixture that is made up of two V-blocks that cradle a solid steel shaft onto which the bearing race is mounted. The steel shaft is marked to aid in placing the indents and is easily rotated while the fixture is mounted in a standard materials test load frame. To conduct a test, the race fixture is placed on the lower platen of the load frame and held in place with  two C-clamps. The ball pushrod is placed in the hydraulically actuated grip of the upper platen. A small amount of silicone vacuum grease is placed into the hemispherical cavity machined into the lower end of the ball pushrod. The ceramic indenter ball is placed into the cavity and the grease prevents the ball from falling out during unloaded periods of testing. The load frame is computer controlled using the desired test load as the control parameter. To run a test, the ball pushrod is lowered at low speed by the computer until contact with the race is  detected. The test load is then gently applied over approximately a five second ramp-up period and then held for 60 seconds. The load is then removed and the race shaft is rotated to the next desired test  location. Dent locations are spaced equally 12 /g113apart around the circumference of the race. After the dent  experiments are complete, the fixture is removed and placed into an optical surface profilometer for dent depth characterization. 78
Figure 6. Bearing Race-Ball Dent Fixture.  Results and Discussions The single ball versus race static load capacity of a hybrid Si 3N4-60NiTi bearing was evaluated and the  results are shown in Table 2. In the table, the dent depth, dent depth normalized by the ball diameter, the  peak stress, and the mean stress are listed. The normalized dent depth as a function of mean stress is plotted in Figure 7. The dent depth data for the ball-race contact agrees reasonably well with the more generalized ball-flat  plate dent depth data collected previously (Ref. 8). This agreement, despite differences in contact geometry, further support the argument that the Hertz stress relations for contacting bodies can be  confidently applied to bearing design. Using the generally accepted dent depth criteria (dp/D~0.00003 to  0.00010), the data indicate that the damage threshold for 60NiTi is around 2.8 to 3.3 GPa. This is slightly lower than the value estimated from flat plate dent studies but within data scatter and is to be considered in good agreement. Further experiments with additional race sizes are warranted to narrow the stress limit more precisely. For the present, a reasonable stress limit, when designing bearings to withstand high  static load, is 3.1 GPa. In the section that follows, a notional spacecraft mechanism, a small reaction wheel assembly (RWA), will  be assessed to determine how the substitution of 60NiTi for steel in the bearings can affect the ability to withstand the rigorous launch vibration load environment. Through such an assessment it is hoped that  pathways to performance improvement, weight reduction or enhanced robustness will be revealed. 79
Table 2.  Indentation Depth Data Summary [{Si 3N4Ball (8.74-mm dia.) loaded against hardened (HRC 60.5) polished 60NiTi inner race} Inner Race Curvature Radii: ball-path, 4.25 mm; cross-race, 1.27 mm.] Indent load, kgf(lb)Mean stress,  GPa (ksi)Peak stress, GPa (ksi)Dent depth,a /g80/g80m (/g80/g80in.)Dent depth/Ball  dia.,/g117/g11710–4  493 (1000) 2.05 (298) 3. 07 (446) None detected ----- 741 (1500) 2.40 (348) 3.5 2 (511) None detected ----- 988 (2000) 2.59 (375) 3.88 (562) 0.19 (7.5) 0.22 1111 (2250) 2.69 (390) 4.03 (585) 0.12 (4.8) 0.14 1235 (2500) 2.79 (404) 4.18 (606) 0.59 (23.3) 0.68 1358 (2750) 2.88 (417) 4.32 (626) 0.55 (21.8) 0.63 1482 (3000) 2.96 (429) 4.44 (644) 0.80 (31.4) 0.92 1605 (3250) 3.04 (441) 4.53 (661) 1.20 (47.4) 1.37 1729 (3500) 3.12 (452) 4.68 (678) 1.84 (72.6) 2.11 1852 (3750) 3.19 (463) 4.79 (694) 1.76 (69.3) 2.01 1976 (4000) 3.26 (473) 4.89 (709) 2.53 (99.8) 2.89 2100 (4250) 3.32 (482) 4.99 (723) 3.68 (145) 4.21 2223 (4500) 3.39 (492) 5.08 (737) 4.01 (158) 4.59 2346 (4750) 3.45 (500) 5.18 (751) 5.53 (218) 6.33 2470 (5000) 3.51 (509) 5.27 (764) 4.62 (182) 5.29 aTypical data scatter range /g1145 percent.  Figure 7.  Normalized Dent Depth Ve rsus Mean Hertz Contact Stress for  Inner Race of a 50-mm 60NiTi Bearing Loaded by a Si 3N4 Indenter Ball  (8.74-mm dia.). 80
Typical steady acceleration forces for rocket launch are 1 to 3 g’s. Thus for the RWA example being  considered, a 5-kg wheel, the maximum applied rotor load is 17.1 g’s (14.1 + 3). Figure 8. Typical Reaction Wheel Assembly.  Figure Based Upon Honeywell Corporation  Model HR 0610 Design (Ref. 11). 5-kg Wheel S upported on Four R4 Ba ll Bearings With  Standard (Catalogue Values) Design (Ref. 12). Table 3. Reaction Wheel Assembly Bearing Configurations Assessed [Ball: 8.74-mm dia., Inner Race Curvature Radii: ball-path, 4.25 mm; cross-race, 1.27 mm.] Configuration  no.Ball material Race material Limiting contact  Stress,aGPa (ksi)Single ball-race load  limit, N (lb) I 440C 440C 2.5 (350) 196 (44) II Si 3N4 440C 2.5 (350) 138 (31) III 60NiTi 440C 2.5 (350) 463 (104) IV 60NiTi 60NiTi 2.5 (350) 846 (190) VS i 3N4 60NiTi 2.5 (350) 374 (84) VI 60NiTi 60NiTi 3.1 (450) 1780 (400)c VII Si 3N4 60NiTi 3.1 (450) 801 (180) VIIIbREX20 REX20 3.8 (550) 587 (132) IX Si 3N4 REX20 3.8 (550) 467 (105) aMean Hertz contact stress.  bREX20 properties: Young’s Modulus (E): 234 GPa; Poisson’s Ratio ( /g81): 0.30.  cHertz calculations may be invalid due to excessively deformed geometry. Table 3 shows the bearing material configurations considered along with the bearing geometry  assumptions and other pertinent details needed for the calculations. The load capacity of the RWA bearing utilizing one of the configurations shown in Table 3 is determined  by combining Hertz stress calculations with generally accepted bearing design relations that correlate ball-race contact stress to radial load capacity. For the Hertz calculations, an automated spreadsheet is  used in which the contact material properties, geometry and single ball-race load is input. The  82
spreadsheet provides the mean and peak stress along with the contact area using the stress  approximation equations developed by Antoine et al. (Ref. 13). The last column in Table 3 gives the single ball-race static load capacity for the nine materials and stress limit configurations considered. To relate the single ball-race load capacity to radial load capacity of the R4 bearing, the relationships  presented by Derner and Pfaffenberger and shown below are employed (Ref. 14):  F R= 9Q max/5 (1) In Equation (1), F Ris the bearing load capacity and Q maxis the single ball load capacity. The R4 bearing  radial load capacity is then multiplied by four to account for the use of two duplex pairs of bearings in the RWA to get the overall wheel load capacity. This value is compared to the launch forces on the wheel to  determine launch design margins for the various configurations. The calculation results are shown in Table 4.  Table 4. Reaction Wheel  Launch Load Capacity [5-kg wheel mass supported on two duplex pairs of R4 bearings.] Configuration  no. Ball material Race materialShaft load  capacity, kN (lb)RWA load  capacity, g I 440C 440C 1.4 (316) 28.6 II Si 3N4 440C 1.0 (223) 20.2 III 60NiTi 440C 3.3 (748) 67.9 IV 60NiTi 60NiTi 6.1 (1368) 124.4 VS i 3N4 60NiTi 2.7 (604) 54.8 VI 60NiTi 60NiTia12.8 (2880)a261.2 VII Si 3N4 60NiTi 5.8 (1296) 118 VIII REX20 REX20 4.2 (950) 86.2 IX Si 3N4 REX20 3.4 (756) 68.5 aLoad beyond yield strength of RWA shaft. The load capacity results reveal some interesting insights. The two conventional bearing configurations (I-440C/440C and II-Si 3N4/440C) have load capacity values near the g force requirements for rocket  launch with additional margin to accommodate bearing preload and other forces neglected in the present  simplified analyses. Another insight is that the use of a high performance material like REX20 offers a  significant improvement in the load capacity, about 3-fold over the 440C bearing. It is this very  improvement that drove the development of such high-carbide tool steels for heavy reaction wheels  despite the poor corrosion resistance and higher production costs of REX20 (Ref. 15).  Finally, the estimated shaft load capacity for the all superelastic bearing configuration (VI) shows a value  well beyond the bending load capability of an RWA wheel shaft sized for the bore of R4 bearings. This revelation suggests that in such an application, a superelastic bearing is essentially shockproof in that the  bearing capability is beyond any load the rest of the instrument could withstand. Of course, designing a  space mechanism with load capacity far in excess of requirements is unrealistic. Excess launch load capacity margins do not enhance system performance. However, the availability of more robust bearing materials, like 60NiTi may enable the use of fewer and smaller bearings thus reducing weight, power consumption, and potentially cost. In addition, the enhanced corrosion resistance and non-magnetic  properties of the NiTi alloys could enable mechanism designs with new capabilities. 83
8.C. DellaCorte, L.E. Moore, III, and J.C. Clifton: “The Effects of Indenter Ball Radius on the Static  Load Capacity of the Superelastic 60NiTi for Rolling Element Bearings,” NASA/TM—2014-216627,2014. To be published. 9.C. DellaCorte and W.A. Wozniak: “Design and Manufacturing Considerations for Shockproof and  Corrosion-Immune Superelastic Nickel-Titanium Bearings for a Space Station Application,” NASA/ TM—2012-216015 and Presented at the 41 stAerospace Mechanisms Symposium, Pasadena, CA,  May 2012. 10. B.J. Hamrock, S.R. Schmid and B.O. Jacobson: Fundamentals of Fluid Film Lubrication , 2nd Edition,  Chapter 21, Marcel-Dekker, 2004, NY.  11. Data Sheet: Honeywell Model HR 0610 Reaction Wheel, Document #DFOISR:03-S-1924,  December 2003, Honeywell Corporation, Phoenix, AZ. 12. Barden Super Precision Ball Bearings Catalogue, D/SPC/1/USA/113/T, Danbury, CT, 2013. 13. J.F. Antoine, G. Abba, C. Visa, and C. Sauvey: “Approximate Analytical Model for Hertzian Elliptical  Contact Problems,” ASME Journal of Tribology, volume 128, issue 3, pp. 660-664, March 2006. 14. W.J. Derner, and E.E. Pfaffenberger: “Rolling Element Bearings,” in CRC Handbook of Lubrication:  Theory and Practice of Tribology-volume II-Theory & Design , edited by E.R. Booser, pp. 495-537,  1983. 15. W. Park, M.R. Hilton, P.C. Ward, G.W. Henderson, A. R. Leveille, D.E. McClintock, and D.W. Smith,  “Microstructure, Fatigue Life and Load Capacity of PM Tool Steel REX20 for bearing Applications,” Lubrication Engineering, volume 55, number 6, pp. 20-30, 1999. 85
to the cylinder, and thereby the load between the contacting rollers, is adjusted by a hand-operated valve  (open-loop control). A turbomolecular pump assisted by a scroll pump provides vacuum in the test  chamber. The typical condition in the test chamber is a pressure of about 4x10-5Pa (3x10-7torr). The most  prevalent remaining constituent in the chamber during testing is water vapor as determined by residual gas analyzer [6]. The residual gas analyzer was used regularly to monitor the test chamber environment with no irregular results. (a) (b) Figure 3.  Schematic views of the vacuum roller ri g. (a)  Schematic, side view. (b)  Schematic,  overhead view with shaft misa lignment depicted and exaggerated. A set of sensors on the test apparatus monitors the test conditions. The outputs of the analog sensors  were digitized at a rate of up to 0.66 Hertz. The misalignment of the driving roller shaft and driven roller  shaft is depicted in an exaggerated manner in Figure 3. The misalignment is measured via a linear  variable differential transformer (LVDT) attached to the bedplate. The LVDT tip moves against a stop onthe turntable. To establish the aligned condition, special tooling blocks were machined to position the roller mounting surfaces as parallel to each other. The precision of this method for aligning is within 0.08  degree. The angular displacement from the aligned condition was calibrated by mounting a laser light source on the moving shaft at the roller mounting location and directing the light onto a paper placed at a known radial distance from the center of the turntable. The movement of the laser light was marked on the paper, and the distance between the points used to calculate the angular movement of the turntable and thereby relate this motion to the LVDT sensor output (voltage).The torque on the output shaft is  monitored by a strain-gage torquemeter of 22 N-m (200 in lb) torque capacity. Calibration was done in  place using deadweights on a torque arm. Figure 4 provides a schematic of the test roller setup labeled  with some of the nomenclature used herein. The load pressing the rollers together is termed the “normal ferrofluid seals and bearingstorque/force sensor output drive  motorpivot point for drive motor  plate relative to bed plate input drive motor bedplatehydraulic cylinder“normal load”  force sensorvac chamber turntable bedplatebellowsmagnetic clutch ferrofluid seal and bearingtorque/force sensor output drive motordrive motor bedplateturntable on bearingload cell for “tangential” force, mounted to turntable magnetic clutch pivot point of drive motor plate relative to turntable 89
Table 2 – Comparison of the Test Rig and Mechanism Contact and Durability Metrics VRRmechanism steel rollerTi6Al4V, two  rollers, same trackTi6Al4V, one roller  only in track duration, contact passes 90,000 288,000 192,000 96,000 /g90/g72/g68/g85/g3/g76/g81/g71/g72/g91/g3/g11/g51/g68/g194/g80/g18/g20/g19/g40/g25/g12/g3a86 120 80 40 pressure, max (MPa) 460 330 rolling speed (mm/sec) 37.2 13.3 contact passing time (sec) 0.005 0.009 a  wear index defined as product of mean of Hertz pressure /g152sliding distance /g152number of contact  passes for 0.9-degree misalignment angle for both system per the approximate sliding  distance concept of Figure 5. Selections of Alternative Materials and Configurations

A set of potential alternative materials were selected based on study of the literature, discussions with  experts in the field, and implementation constraints. Liquid lubrication was not considered because of the extreme cold temperature of this application. Two review papers provide direct guidance and a substantial listing of relevant solid lubrication information [13,14]. Five different fundamental alternative  material candidate solutions were selected for evaluation, as follows:  1. Use of polyimide rollers rather than steel rollers. Such material has been used for space  applications and has been proposed for use as traction rollers. The experiment made use of areadily available VRR roller made from an unfilled base polyimide resin [15].  2. Use of a bonded, PTFE-type solid lubricant coating on the steel roller. The binder was an inorganic type with highest available strength.  3. Use of an alternative surface treatment to the currently employed anodizing treatment of the  Ti6Al4V material, with intent to provide an adherent, hard, wear resistant layer of oxides. The selected treatment of the Ti6Al4V was a plasma electrolytic oxidation process [16]. 4. Use of metal-doped diamond-like carbon coatings applied one or both surfaces to provide an  adherent layer of high hardness and toughness. Some success has been reported for this approach for prevention of cold welding of Ti6Al4V in vacuum [17]. Such an approach has also been reported to improve durability of oil-lubricated bearings under certain operating conditions  [18]. Two such coatings were selected, one a (TiC)aC:H type and one a (WC)aC:H type. The  coatings were applied by physical vapor deposition. 5. Use of a molybdenum-disulfide (MoS 2) type sputtered solid lubricant film applied to one or both  contacting surfaces. The selected film is a nanocomposite coating with composition including antimony oxide and gold. Such compositions have been studied by others for use in vacuum conditions [19]. MoS 2has an extensive heritage as an effective solid lubricant for cold vacuum  conditions. The coating thickness was intended to be 1 micrometer.  The material configurations selected for testing evolved as test results were obtained. Ten configurations  were tested as listed in Table 3. 92
(a) (b) (c) (d) (e)  (f) (g)(h)  Figure 6. Conditions of rollers at end of test. (a) Config. 1, annealed 440F steel. (b) Config. 2, hardened 440C steel.   (c) Config. 3, polyimide upper roller.  (d) Config. 4, bonded film.   (e) Config. 5, PEO of Ti6AL4V.  (f) Config. 6, lower roller coated with (WC)aC:H   (g) Config. 7, both rollers  coated with (TiC)aC:H.  (h) Config. 8, upper roller coated with nanocomposite MoS 2 95
(a) (b) Figure 7 – Conditions of rollers at end of test fr om an unintended but inte resting full sliding test. (a) Upper MoS2 nanocomposite coated roller.  (b) Uncoated lower roller, stalled member. /g38/g82/g81/g73/g76/g74/g88/g85/g68/g87/g76/g82/g81/g3/g20/g19/g3/g177/g3/g53/g72/g86/g88/g79/g87/g86/g29/g3 The tenth and final configuration for discussion was almost identical to  Configuration 9, except there was one additional processing step by the vendor in preparing Ti6AL4V  rollers for sputtering. The nanocomposite MoS 2coating was present on both upper and lower rollers for  this configuration. It was noticed that the anodized surface of the titanium alloy rollers were quite rough, and roughness of such magnitude was judged as not favorable for minimizing wear. The lower titaniumalloy roller surfaces were smoothed by directing ultrasonic energy to the surfaces using a hand held probe. Aside from this added “smoothing” step for the anodized Ti6Al4V rollers that was done just prior to application of the coating, Configurations 9 and 10 were otherwise identical. Six tests were conducted forthis configuration to more than 87,800 cycles for each test. Four of the tests provided excellent wear  protection. Two of the tests, while slightly lesser performing, still provided very good wear protection. For  one of the two lesser performing tests, the coating thickness was about one-half of the intended thickness  because of a temporary problem with the coating process. The thinner film was likely at least partially responsible for the less effective wear protection. .For the other lesser-performing test, a narrow wear track developed after about 60,000 cycles covering about one-third of the contact width. The photograph of Figure 9 shows typical end-of-test conditions of the rollers for the four tests that provided excellent  wear protection. Profilometer inspections showed essentially no wear, to the precision that could be measured by the  stylus inspection technique, for both rollers. Typical lower roller inspections at start and end of test are  provided in Figure 9(b-c). The undetectable wear on the lower roller is in quite contrast to the behavior of Configuration 9, per Figure 8(c) whereby a distinct wear valley was produced. It is interesting that the  asperity features within the wear track of Figure 8(c) are very similar in shape and magnitude to the asperity features of the ultrasonically-smoothed and coated, but untested, roller of Figure 9(b).The many  bright particles of debris easily seen without magnification on the debris pan during tests of Configuration 9 were absent during all tests of Configuration 10. Comparing the profilometer inspections of the lower  rollers before test shows the effect of the ultrasonic smoothing process. Without the ultrasonic smoothing,  the surface has asperities with high slopes as had resulted from the anodizing process. The asperity  peaks are removed and/or deformed by the ultrasonic smoothing process [compare Figure 8(b) to Figure 9(b)] providing a more favorable surface for avoiding wear. 98
(b)   (a)   (c)  Figure 8 – Data from Configuration 9. (a) Conditions of rollers, end of test. (b) Profilometer  inspection, lower roller, start of test. (c) Prof ilometer inspection, lowe r roller, end of test.   (b)   (a)    (c)  Figure 9 – Data from Configurati on 10. (a) Conditions of rollers, end of test. (b) Profilometer  inspection , lower roller , start of test. (c) Profilometer ins pection , lower roller , end of test. inspection direction inspection direction 99