ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
bearings. Although this may be an advantage for continuously rotating higher speed mechanisms acting  as an additional lubricant  Figure 11. Post life test bearing  Figure 12. Post life test bearing raceway Figure 13. Post life test b6amlllY mabcway  CRISM Bearing Anomalies  Bearina Ball Spacing  The APL Space Department has used different sizes of thin-section Teflon-toroid-spaced bearings on  several programs and tests. One feature that we have always noticed with these types of bearings is the  potential for a “large” gap to occur between a ball and toroid (Figure 14).  We attribute both the high running torque and the poor feel to the non-uniform ball-toroid spacing. We  believe that non-uniform ball-toroid spacing can result in some various amounts of pressure between all  the toroids and balls. This can result in unpredictable friction between the ball and toroid causing an  increase in the running torque. It may also result in a stick-slip situation between the ball and toroid  resulting in torque spikes or non-smooth rotation. Additionally, it could lead to rapid wear of the toroid.  The toroid wear particles could wind up in the raceways as contamination, also causing anomalous  19
torque spikes. Grease and oil lubricated bearings may exhibit the same problems as the CRISM dry film  lubricated bearings to a far less noticeable degree due to the grease or oil between the ball and toroid.  The evidence for this conclusion with the CRlSM bearings is two fold:  1. The significant amount of Teflon contamination found in the bearings following the run-in  procedure described in Appendix 1  2. The drastic change in torque level and smoothness following the high-pressure air blow through  cleaning procedure described in Appendix 1  A  Figure 14. Thin section bearings, same size, different manufacturer  ReDroducible Torque SDikes  Another type of torque spike was also noticed with the CRlSM bearings that could be easily reproduced  based on the operation of the spindle. During the run-in procedure in Appendix 1, unidirectional rotation  resulted in extremely smooth running torque. However, CRISM was intended to oscillate do”. A  reproducible torque spike occurred following a change in direction of rotation. Rotating the spindle  backwards and forwards, sometimes referred to as “safe cracking”, torque spikes, equal to or greater than  the nominal running torque, would result within several degrees of rotation following the reversal. A ball  can never remain rolling between surfaces that form an angle to each other5. All angular-contact ball  bearings create an angle between the two raceways. Therefore, as the bearings rotate, the balls produce  a gyroscopic motion in addition to rolling. Pressure or drag friction between the toroid and the adjacent  balls also seems to deflect the toroids based on the direction of rotation. Thus, the gyroscopic motion of  the balls and pressure between toroids and adjacent balls appears to cause to cause the toroids to align  themselves based on the direction of rotation. This behavior appears to be attributable to the vast majority  of toroids aligning themselves (Figure 15). When the spindle reverses, as do the balls, the toroids flip and  align themselves in the opposite direction (Figure 16). As the toroids flip and align themselves in the  opposite, direction, a torque spike resulted. This effect was significantly reduced following the cleaning  procedure. Once the balls and toroids became more evenly spaced, the drag friction between the toroid  and adjacent balls was reduced, thus reducing the ability of the toroids to align themselves. This was also  noticed following the Christmas holiday. As the bearings sat over the holiday, pressure between the balls  and toroids either slightly re-spaced the balls or caused the Teflon to cold flow resulting in bearings the  felt much better after having sat for an extended period. However, following another run-in, the bearings  quickly resorted to their pre-holiday behavior.  20
L Figure 15. Toroids in “down” position  Unfortunately, we never photodocumented this effect, thus, it was extremely difficult to find pictures of the  flight hardware demonstrating this effect clearly. Figure 15 clearly shows all the toroids uniformly aligned  in the “down” position. Figure 16 shows most of the toroids in aligned in the “up” position, including most  of the ones at the top of the bearing. This effect can not be reproduced with a single unmounted bearing.  CRISM Bearing Assembly Procedure  The CRlSM bearings were removed from the manufacturer’s packaging and visually inspected to verify  that they conformed to the documentation and were marked properly (Figure 17). Once we were sure that  the bearings were correctly marked and free of contamination, they were assembled as follows:  1. The bearings and spacers were stacked and aligned per drawing and documentation (Figure 18)  2. The bearings and spacers were placed in an assembly fixture (Figure 19) specifically designed to  keep them aligned during assembly  3. The bearing retainer was placed on the assembly (Figure 20)  4. The bearing shaft (Figure 21) was cooled in liquid nitrogen  5. The cooled bearing spindle was quickly assembled in the bearing assembly fixture. Weights were  placed on the assembly to ensure that it remained seated against the top bearing.  6. The bearing assembly was quickly placed in a N2 purged vessel and allowed to equilibrate for 24  hours (Figure 22)  7. The spindle bearing assembly was then visually inspected and checked for “feel”. It was noted  that there was a significant amount of pressure on the spacers between the two bearings on both  the Engineering Test Unit (ETU) and the flight unit. The ETU and the flight assembly were the first  bearing assemblies with spacers that we had assembled using the liquid nitrogen technique.  Previously, the bearings were duplex pairs without spacers. We were concerned that with the  shaft being significantly cooler than the bearings and spacers, that as it warmed and expanded,  there would be very little pressure between the bearings and spacers or worse, a gap. However,  that never materialized  8. The entire spindle bearing assembly was then cooled in liquid nitrogen  9. The cooled spindle bearing assembly was quickly assembled in the titanium  motor/encoder/bearing housing  10. The assembly was then purged and allowed to equilibrate for 24 hours  11. The assembly was visually inspected and checked for “feel”  12. We were not happy with the “feel” of the flight assembly and took the steps outlined in Appendix 1  There were two main issues with the flight bearing assembly, 1) high running torque; 2) torque spikes that  made the assembly not feel smooth as it rotated. We were confident that contamination was not the  problem from pre- and post-assembly inspections causing either of these problems.  21
Appendix 1  Date Operation  12/04/2003 Initial torque measurements Table 2 lists the operations and measurements made to the flight motor/encoder bearing assembly. Most  measurements were an average running torque. Some measurements were a peak torque when there  was a significant torque spike to the running torque. On several tests where many measurements were  made, an average value is reported and denoted by and Avg. in the Torque column.  Torque  mN-m (oz-in)  111.6 (15.8) Table 2. Bearing Assembly Operations  12/16/2003 11 1.6 (1 5.8) nominal  190.7 (27.0) peak  Run-in Q60 RPM, 19 minutes (1 140 Revolutions) 143.0 (20.25) Avg.  Run-in Q60 RPM, 25 minutes (1500 Revolutions) 381.3 (54.0)  Cool off 254.2 (36.0)  12/24/2003-  1 /4/2004 Loosen retaining nut  Run-in Q60 RPM, 20 minutes (1200 Revolutions)  Tighten retaining nut, 17.0 N-m (1 2.5 ft-lb)  Stored in N2 purged container in cleanroom 158.9 (22.5)  190.7 (27.0)  1 /5/2004  Changed torque measuring method for all future  measurements.  Peak torque measurements  ~ Loosen retaining nut and re-measure torques  Set-1 238.7 (33.8) Avg.  Set-2 317.8 (45.0) Avg  1 /6/2004 Run-in Q60 RPM, 20 minutes (1200 Revolutions)  Peak torque measurements  The best these bearings have ever “felt”  Removed retaining nut, rotor, and spacer 148.3 (21 .O) Avg.
Acknowledgements  This work was performed at Johns Hopkins University Applied Physics Laboratory under a contract with  the National Aeronautics and Space Administration. The authors would like to thank all those who helped  design, fabricate, integrate and test the CRISM instrument. Reference herein to any specific commercial  product, process, or service by trade name, trademark, manufacturer, or otherwise does not constitute or  imply its endorsement by the United States Government or the Johns Hopkins University Applied Physics  Laboratory, Laurel, MD.  References  1. S. Murchie et. al., “CRISM (Compact Reconnaissance Imaging Spectrometer for Mars) on MRO (Mars  Reconnaissance Orbiter)”, Instruments, Science, and Methods for Geospace and Planetary Remote  Sensing, SPlE Vol. 5660, pp. 66-77.  2. Bugby, D., J. Garzon, 8. Marland, 6. Stouffer, D. Mehoke. M. Fasold, “Cryogenic Diode Heat Pipe  System for Cryocooler Redundancy,” SPlE Optics and Photonics Conference, Cryogenic Optical Systems  and Instruments XI, San Diego, CA, 31 July -4 August, 2005.  3. J. Lees, E. Schaefer, “Design and Testing of the CRISP Tracking Mirror Cover and Release  Mechanism”, Proceeding of the 36‘h Aerospace Mechanisms Symposium, Glenn Research Center, April,  2002, NASNCP-2002-211506, pp.63-76.  4. Esplin, Roy. A satellite-based multichannel infrared radiometer to sound the atmosphere (SABER). In  Optical Remote Sensing of the Atmosphere held in Salt Lake City, Utah, 5-9 February 1995. pp. 130-1 32.  5. Harris, Rollina Bearina Analvsis, 3rd Edition, John Wiley & Sons, Inc., pp.451-452.  24
Gear Teeth Particles and Bearing Failures  William H. Greenwood* and Jeffrey G. Dabling*  Abstract  Torque is transmitted from rotary solenoids to rotate drive arms that advance a ratchet wheel as part of a  safety mechanism in missile warheads. The small volume constraint led to single gear teeth to transmit  the torque from the rotary solenoid. High contact forces and compliant gear teeth caused many fine  particles to be generated at the rubbing surfaces of the gear teeth. The particles were pulled into the ball  bearings of the adjacent solenoids causing early failure while having no ill effect on the ball bearings of the  drive arms. A temporary solution of custom plastic shields allowed the prototype units to proceed to  environmental and flight tests. A subsequent build replaced the gear teeth with ball bearing followers.  Introduction  Missile warheads and bombs usually have safety mechanisms [l ] to prevent unintended explosions in the  event of accidents. These safety mechanisms are small and made of stainless steels for high temperature  integrity in the event of accidents involving jet fuel or propellant fires. The safety mechanisms often have  ratchet-type wheels, and are usually driven by rotary solenoids (Figure 1).  Figure -  Mounting plate, ratchet wheel, and rotary solenoid  * Sandia National Laboratories, Albuquerque, New Mexico  Proceedings of the 3dh Aerospace Mechanisms Symposium, Langley Research Center, May 17- 19,2006  25
A new generation of safety mechanisms is nearing production. The safety mechanism of Figure 1 is  among the new generation and transmits torque from the rotary solenoid to the drive arms by way of  gears. Due to the limited stroke of the solenoid and the volume constraint, the gears are actually single  gear teeth cut as part of the solenoid rotor and drive arm (Figure 2).  Figure 2. Rotary solenoid and single gear teeth  Earlier generations of these safety mechanisms often used full gears to transmit torque from the ratchet  type wheels to other shafts for thousands of ratchet wheel cycles. One life test of the first generation of  safety mechanisms involved 10,000 ratchet wheel cycles; the main gear had teeth that were no longer  involutes but appeared triangular. Much wear debris was adjacent to the gear but did not migrate to ball  bearings or cause operational failure.  The first units of the new intent safety mechanism were operated in January 2003 and failed to operate  after only a dozen to a few hundred ratchet wheel cycles, as shown in Table 1. A failure to operate at such  low number of cycles was unexpected. The original units had molybdenum disulfide (MoS2) applied to the  gear teeth, but had no lubrication on the ball bearings of the solenoid or the drive arms. Molybdenum  disulfide is used since the mechanisms must sit for decades unused and then be called upon to function.  The safety mechanisms are usually hermetic units and have a nitrogen and helium atmosphere, although  the ratchet cycle tests in this report were all performed in the laboratory atmosphere as part of early  qualification. The ball bearings are generally lubricated but a new process for the bearings was not yet  ready, so the early ratchet wheel cycle tests were performed with dry (non-lubricated) ball bearings.  26
al  Figure 3. Gear tooth wear zone has a powdered layer.  The solenoids from failed units were rotated manually and some roughness in rotation was noted.  However, a quasi-static torque test of the solenoid at 16 and 22 volts showed nearly the same torque as  measured before the installation into the ratchet wheel assembly, but the measured friction at zero volts  was somewhat higher. The solenoid bearings in failed units were visually examined at 1 Ox magnification,  and the balls in the bearings in the inboard rotor had a roughened or a worn surface (Figure 4). The  solenoid ball bearings were replaced, and the solenoids were re-assembled into their failed ratchet wheel  units. The rebuilt units were then measured for minimum operate voltage and found to be 15 volts, the  same as new units. The parts that degraded as the ratchet wheel units were operated were the solenoid  bearings. The rebuilt units (with new solenoid bearings) were then cycle tested again and failed at similar  cycles as those of Table 1.  The solenoid bearings could be degrading due to high dynamic loads or from particles from the gear teeth  entering the bearings during operations. The axial load from the magnetic forces when the coil is  energized was calculated to be 71 N (1 6 Ibf) compared to the dynamic load rating of 11 6 N (26 Ibf). The  small oscillatory solenoid stroke of 0.2 rad (12 degrees) is a more severe service than continuous  rotations since only a small portion of the balls see the load repeatedly. Therefore, the high axial magnetic  force could be degrading the unlubricated bearings in such a few cycles. However, the innermost bearings  of the rotors support all of the axial forces; whereas the failed units showed bearing wear to be most  severe in the inboard rotor and approximately equal from the innermost and outermost bearings of the  inboard rotor. The rapid bearing degradation did not appear to be caused by the relatively high dynamic  axial load.  28
,A ri F- A  PA  Figure 4. Solenoid ball bearings before (left) and after (right) life cycle test.  When ball bearings fail from high loads, the initial signs are many small pits as bearing material spalls,  then the ejecta from the pits are rolled into the bearing surfaces that cause more localized high loads and  accelerated degradation and eventually a non-rotating (seized) bearing. As the bearings were visually  scrutinized, many patches or scabs were observed and some indentations but not pits. The scabs  appeared to be a particle or perhaps several particles that are rolled into the balls as the bearing is  operated. The scabs were foreign particles and not bearing ejecta. The parts that degraded and caused  the early wheel cycle failure were the solenoid bearings; the cause of the bearing degradation was the  particles from the gear teeth. The problem resolution was to sharply reduce the particles generated at the  gear teeth.  Attempts to Reduce the Gear Teeth Particles  The gear teeth were lubricated with a proprietary molybdenum disulfide process. A tribological  investigation showed some alumina particles imbedded in the surface where the solid lubricant was  applied. The lubricant applicator did indeed prepare the surfaces with alumina beads. A possible cause  was identified; some of the alumina beads from the surface preparation were imbedded into the gear  surface and caused grinding at the mating gear surfaces and generated the great amount of metal  particles.  Gear teeth surfaces were cleaned of the MoS2 and alumina beads and were then life cycle tested, but the  same failure to operate after a few hundred ratchet wheel cycles was observed (Table 2). The post-  mortem inspection again showed many metal particles at the gear teeth wear zones and scattered on the  nearby mounting plate and stop pins. The inboard solenoid rotor had much rougher manual rotation than  the outboard rotor. The wear zone on the inboard rotor gear tooth was broader at the base as shown in  Figure 3, rather than a more parallel wear zone observed in the outboard rotor gear tooth.  29
Table 2. Ratchet wheel rotations to failure, no MoS2 on gears, no MoS2 on bearings  CP008  CP009 12 Feb 2003 180 Rough manual rotation  14 Feb 2003 260 Rough manual rotation  Several units with the wear resistant outboard gear were tested with the same disappointing result: the  units failed to operate after a few ratchet wheel cycles, and there was a thick powdered layer at the wear  zone on the gear teeth (Table 3 and Figure 5). The wear zone on the new outboard gear appeared to be  even heavier than the wear zone on the trials from Tables 1 and 2. The rotors spun roughly after the life  test. The bearings were observed to have a roughened surface on the inboard rotor, but only slightly  roughened on the outboard or stator bearings. The long stringers observed in Figure 5 on the edge of the  gear are composed of many fine particles that are magnetized. These magnetized stringers appear at the  edges of other parts as well as the gear teeth. A schematic of the solenoid is shown in Figure 6. The  solenoid bearings from left to right will be numbered 1 through 4 in many of the descriptions to follow. The  outboard rotor bearings are 1 and 2 while the inboard rotor bearings are 3 and 4.  Figure 5. Wear particles generated due to a Nitronic 60 gear tooth interface  30
Table 3. Ratchet wheel rotations to failure, Nitronic 60 gear, no MoS2 on gear or bearings  Date Wheel cycles p  CP011 20 Feb 2003 72 Bearings 3 and 4 bad, bearings 1 and 2 fair  CP037 20 Feb 2003 280 Bearings 3 and 4 bad, bearings 1 and 2 fair  CP034 21 Feb2003 220 Bearing 3 very bad  CP045  CP054  CP021 Outboard rotor Outboard coil  -- 27 Feb 2003 110 I  27 Feb 2003 90 Inboard rotor rotates very roughly  03 Mar 2003 51 8 Inboard rotor bearing seized Stator  24.9 mm  L Inboard coil  Inboard rotor  I_- 19.2mm -4 Inboarc \  Ball bearings i gear  Figure 6. Double rotary solenoid; outboard and inboard rotors move independently.  A new trial was performed with MoS2 to the applied gear teeth, but with the surfaces prepared by acid  etching rather than alumina bead blasting as in the original condition of Table 1. The desire was a drastic  reduction of wear particles at the gear teeth. Again, low ratchet wheel cycles to failure were measured and  a heavy powdered layer at the gear teeth was again observed (Table 4). The third replication that failed at  wheel cycle 518 was encouraging, but the two replications that failed at low cycles indicate that the  condition of Table 4 is similar to the first three conditions. The unit that failed at 51 8 cycles was examined  carefully in the post-mortem, but resulted in no clues regarding its much longer operating life.  Table 4. Ratchet wheel rotations to failure, MoS2 on gears without grit blasting  31
A Temporary Solution - Bearing Shields  As our early efforts to decrease particle generation at the gear teeth were unsuccessful (Tables 2-4), we  looked for a quick solution; so that environmental and flight tests could continue. A shielded bearing was  suggested as a way to prevent the gear teeth particles from entering the solenoid bearings. However,  there is not room for standard shielded bearings. The inboard rotor is 3.8-mm thick; the original bearings  are 1.6 mm thick, which leaves a 0.6-mm gap between the ball bearings. A shielded bearing of the same  inside and outside diameters has a thickness of 2.4 mm that for a pair of bearings is wider than the rotor.  The temporary solution was to make thin washers, 50 microns (.002) thick, out of acetate plastic that had  the same inside and outside diameters as the bearings and were placed next to the bearings of the  inboard rotor. Plastic was used since they could be made quickly. One shield was placed on the inboard  rotor bearing next to the outboard gear (bearing 4) as shown in Figure 7, and the other shield was placed  between the inboard rotor bearing (bearing 3) and the stator bearing (bearing 2).  These shields were placed in test units with no MoS2 on the gears or bearings and had encouraging  results as shown in Table 5. Finally, most units run at least 500 wheel cycles. The initial solenoid torque  tests showed no increase in friction due to addition of the plastic shields. After the life tests, the shields  showed a darkened annulus where the bearings were sliding, but the plastic was not grooved. A workable  solution appeared possible. The shields appeared to be protecting bearings 3 and 4 since tested units  now had some bad bearings at position 1 with some good bearings at position 4.  -- I  Figure 7. Plastic shields were used to keep particles out of the solenoid bearings  32
Why so many gear teeth particles?  The tribological investigation continued to evaluate material pairs or lubrication that would decrease the  gear particles. Some standard pin on disk tests at a high contact stress showed long life of the MoS2  lubricant and little particle generation. This caused us to calculate the Hertzian contact stress [2] at the  gear teeth. The contact stress was 753 MPa (109 ksi). This exceeds the yield strength of the Hiperco 50A,  365 MPa (53 ksi), used for the inboard rotor, the wear resistant stainless steel, Nitronic 60, 414 MPa (60  ksi), used for the outboard gear in one of the trials, and the drive arms PH 13-8 Mol condition H1150 , 620  MPa (90 ksi). The contact stress was undoubtedly higher on the single cantilevered gear tooth of the  inboard rotor since it deflects during loading and the wear zone on the tooth was more triangular than  rectangular. Figure 3 shows just such a wear zone on the drive arm gear tooth that mates with the inboard  rotor gear tooth. The single gear tooth was analyzed for bending stresses at the base during the initial  design but the Hertzian contact stresses were not calculated for the contact load at the gear teeth. The  high contact stresses mean plastic deformation at the load zones, non-rolling contact and resultant  particle generation.  Long term solution - eliminate the gear teeth  The plastic shield was a fix that permitted further testing of the first build. However, the many particles  were undesirable and created a potential for other problems. A long-term solution was desired where few  particles were generated. Increasing the gear pitch or the length of the gear teeth to decrease the contact  stress by a factor of two or three was not possible due to the volume constraints. In addition, the single  gear tooth on the inner rotor and the drive arms made these expensive parts and difficult for part  34