ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Conclusion The stated objectives of the bushing test series were to characterize the friction and wear performance of  several candidate bushing/coating materials in a flight-like application and evaluate the behavior of these materials over the MSL operational temperature range. Both of these objectives were met during the  riction torque in the pivots was deemed  ove valuable for future Mars Surface Systems considering  e use of bushings in highly loaded pivots. lifornia Institute  f Technology, under a contract with the National Aeronautics and Space Administration.  ank the following individuals and their respective teams for their contributions  h Center: Jaime Waydo, Jet Propulsion Laboratory:   MSL Mobility Mechanical Team Lead  ust 2006)  . 650.  6) 5.  Lince, Jeffrey, “SEM of MSL Lube Test Results”, PowerPoint Presentation (August 2006) testing and subsequent data analysis. Although all test samples survived the modified 4x life test there were clear winners with respect to  friction torque and wear. The MoS 2-based bonded film coated samples exhibited the best overall low  friction behavior while the polyamide-imide samples had the least amount of wear among all bushing  systems tested. Ultimately, the team selected the phosphate-like binder MoS 2-based bonded film /  stainless steel bushing system for flight usage as minimizing the fmore important than minimizing pivot wear during the mission. The bushing test series has provided valuable friction and wear data on a variety of candidate bushing  materials and lubricants enabling the MSL Mobility Team to make critical design decisions about flight pivots on the vehicle. This data should also pr th Acknowledgements  The research described in this paper was carried out at the Jet Propulsion Laboratory, Ca o The author would like to th to this bushing test effort:  Donald Striebing,  NASA Glenn Research Center:     Phase One Bushing Testing  Dr. Christopher DellaCorte,  NASA Glenn Researc Phase One Bushing Testing Dr. Phillip Abel, NASA Glenn Research Ce   Phase One Bushing Testing nter: Ray Andres, Jet Propulsion Laboratory:    Test Fixture CAD Design Support Dr. Jeffrey Lince, Aerospace Corp:    Sample Wear Analysis  Christopher Voorhees, Jet Propulsion Laboratory:  MSL Rover Mechanical Sys. Engineer References 1.  Striebing, Donald, “MSL Mobility Risk Reduction Bushing Test Series” (Aug 2.  Waydo, Jaime  "Dither Analysis Code." Matlab Script , (September 2005).  3.  Young, W. Roark’s Formulas for Stress and Strain 6 th Ed.  Dallas, McGraw Hill, Inc., ©1989, p 4.  Lince, Jeffrey, "SEM of MSL Lube Test Results", PowerPoint Presentation (November 200 96
the pressure and limit the volume that must be pressurized. The valves also allow the recirculation pumps  to operate to chill the engines prior to launch. In order to prevent overpressure of the feedlines, the valve  contains a reverse flow relief valve and a visor liftoff mechanism.   The pre-valves are pneumatically actuated, bi-stable, two-position valves. The valve has a half-moonshaped inconel visor that seals against a Kel-F seat in the closed position. A photo of the valve between  the open and closed positions is shown in Figure 1. To open the valve, helium is supplied to an actuator that rotates the visor out of the propellant flow path. When the valve opens, two Belleville spring-loaded  mechanisms hold a small roller in detent grooves on each side of the visor to ensure the visor remains in  the fully open position. These detent mechanisms serve as a mechanical latch to hold the valve in the open position to prevent an unwanted closure of t he valve during engine operat ion. Because the valve is  bi-stable, the detent mechanisms add redundancy to the valve. Helium is again supplied to the opposite  side of the valve actuator piston to close the valve.    Figure 1:  Space Shuttle Main Propulsion System Pre-Valve  Pre-valve Detent Mechanism Description  The detent mechanisms are composed of several piece parts as shown in the cross sectional and  exploded views in Figures 2 and 3. The assembly contains a roller, pin, and follower that allow the mechanism to translate across the visor. The assembly also contains a stack of Belleville springs and  spacers. The springs allow for vertical movement of the roller. The spacers protect the springs from  rubbing on one another and allow the springs to invert.    The detent roller is manufactured from LO 2/LH 2 compatible Vespel SP-21. As the valve opens, the part of  the visor known as the ramp comes into contact with the roller compressing the Belleville spring stack in  the detent mechanism. The roller travels along the ramp to the detent groove, shown in Figure 4. Figure 5  shows the valve in the open position with the roller locked in the detent groove. Rotation of the visor itself is controlled by mechanical stops in the actuator that do not allow the visor to rotate more than 90  degrees. The stop also prevents the roller from rolling through the groove and out the other side. As the  roller moves up the ramp, the follower also moves up. This upward motion compresses the Belleville spring stack. As the stack is compressed, the spacers slide within the detent cap guiding the entire  stack’s motion. They also keep the four Bellevilles from contacting each other as each is compressed so  that they can move beyond the point where they invert. The top spacer contacts a lip on the cap and transmits the force of stack deflection through the cap and into the seven bolts holding the mechanism  98
within the valve body. The cap retains a static spring  energized seal in th e valve body that prevents  propellant leakage into the aft compa rtment of the Orbiter. When the valve is commanded closed, the  actuation force generated overcomes the force generated by the roller in the groove and compresses the  spring stack until the roller is forced out of the detent groove. The roller then rolls down the visor ramp until the two parts are no longer in contact with each other. In the closed position, the detent Belleville’s  are only compressed to their in stallation height and the roller and follower hang inside the valve. The  follower has a large land that contacts the re tainer to hold it in place. The retainer serves multiple  purposes:  it prevents the follower and spring stack from falling into the valve and ensures the spring  stack is compressed to the installation height before coming into contact with the visor, the retainer is  keyed within the cap to prevent the mechanism from being installed in a manner where the roller and detent groove would not line up, and the retainer prevents the follower from pivoting or rotating as the  roller moves up the visor ramp.    Figure 2:  Cross-section of the detent mechanism  99
Figure 5:  Roller locked in the detent grove when the valve is in the open position  Detent Mechanism Inspections  During an inspection of OV-104 (Atlantis) after the vehicle’s second flight, one of the detent rollers in the  engine 3 LO 2 pre-valve was found to have a crack through th e full thickness and down one third the  length of the part. The root cause of the crack was not found and the crack was dismissed as a one-time  occurrence. However, as part of the forward plan, the requirement to inspect each roller during every  vehicle’s Orbiter Maintenance Down Period (OMDP) was instituted. During every OMDP, all six LO 2 and  LH2 detent mechanisms are removed from the pre-valves and inspected for wear or damage. Each roller  typically shows a washboard pattern around its circumference that is considered nominal wear. This  pattern is due to the large visor swinging and contacti ng the motionless roller. Be cause the surface of the  visor is relatively rough and the visor moves at a fast rate, it is suspected that the roller does not roll  smoothly across the visor. The slipping of the roller may contribute to the “wash board” wear marks that  are typically seen around the roller.    As the Belleville spring is compressed, the roller is forced against the retainer that results in wear on the  roller. Also, there are typically black indications on the retainer where this contact takes place. The black marks indicate deposits of Vespel material were transferred from the roller surface to the retainer.    Some rollers within the fleet are specially machined to have a smaller diameter in specific locations. The  purpose of this process is to reduce the effect of rough or high spots on that specific valve’s visor. During  inspections after the initial flights of the OV-103 and OV-104, several rollers in the LH 2 and LO 2 systems  were found with rough gouges at specific distances along their lengths. Borescope and visual inspections  concluded these instances of damage were due to wear associated with the high spots on each visor.  Two rollers showing this type of damage are shown in Figure 6. Examples of specially machined rollers with gaps to avoid a high spot on a discrepant visor are shown in Figure 7.    Figure 6:  Rollers Exhibiting Wear due to High or Rough Spots on the Visor  101
Figure 7:  Drawings showing how these rollers are machined to avoid discrepant areas   on valve visors  In addition to roller wear, the retainers also show signs of wear due to contact from the side of the  follower. Deposits of the retainer material have been found on the pin and follower, which further supports the theory that the retainer is being damaged by th ese components. The cap also  shows signs of wear  due to the sliding contact of the four spring spacers. Example of nominal wear associated with spacer  travel in the cap can be seen in Figure 8. All parts of the assembly are inspected thoroughly and replaced if the wear is considered si gnificant. Minor wear to metal parts may be treated with Chem-film which is an  acceptable practice for Shuttle LO 2 and LH 2 systems.  Figure 8:  Cap showing signs of nominal wear due to spacer rubbing  Cracked Roller Anomaly  To date, three cracked rollers have been found through inspection. All 3 rollers were in the OV-104  Engine 3 LO 2 pre-valve. The first cracked roller was found after the second flight of OV-104, which led to  the inspections discussed above. During OMDP for OV-104 after Flight 26, the second cracked roller was  discovered. Following this investigation, every-flight inspections were instituted for the specific Engine 3  LO2 pre-valve. During these inspections, the third cracked roller was discovered. Unlike the prior two  cracked rollers, this roller cracked after only one flight. This raised concerns that a cracked roller that remains in service may degrade further and cause a serious failure in the LO 2 system.  A cracked roller may liberate debris. This debris should be captured by the pre-valve screen, which is a  1000-micron screen downstream of the pre-valve to prevent contamination from entering the engine.  102
Rationale to fly after finding the second cracked roller was based primarily on the conclusion that the  sharp leading edge of the detent groove was the root cause of the cracking. However, after the detent  groove leading edge was returned to print, the third roller cracked after only one flight. The second and  third cracked rollers can be seen in Figure 10. In each case, the crack was observed to run through the LO 2 compatibility testing mark and completely through the thickness of the roller. This can be seen in  Figure 11.  Figure 10:  The second and third cracked rollers  Figure 11:  Crack can be seen through the compatibility testing mark and from the   inner to outer diameters  104
During further examination of all availab le data it was noted that the visor to valve body height, recorded  during the visor mold impression and grinding wo rk, was out of family. This information kicked off a  tolerance stack up analysis. The  third cracked roller kicked-off an effort for a more in-depth investigation  of alternate causes. Because the first two causes that were examined, the dimple and the sharp edge, did not preclude another cracked roller, it was necessary to look for a less obvious cause. Because the third  failure was after only one flight, resources and attention not previously available were given to the  problem.  The out-of-specification measurement of the visor height prompted the tolerance stack-up investigation.  The measured distance of the visor from the valve body was used as a basis for a tolerance stack-up analysis. Actual measurements of the installed detent mechanism components were used, including a  force-deflection curve of the Belleville spring stack set-up. The stack up of the individual components was  subtracted from the visor height to determine if interference was possible. Interference would result in the  spring stack reaching its solid height and imparting unexpected high loads on the roller. It was determined  that combinations of nominal piece parts tolerances could result in deflections that caused the Belleville spring stack to go flat and introduce  the high loads onto the roller.  An illustration of the  mechanism and  the measurements used in the analysis are shown in Figure 12. As a solution, piece parts were carefully  selected to produce an acceptable assembly on the low side of the tolerance band. This should preclude the spring stack from reaching a solid deflection height. Further mitigation was to machine an angle in the  retainer surface that contacts the follower. The surface was angled to 2 degrees based on the marks left  by the roller on the retainer. These marks indicated the angle at which the roller was moving when in contact with the visor. The modified retainer allows additional movement of the follower and prevents  binding at the follower-retainer interface. This reduction in binding also reduces the load being applied to  roller.  Figure 12:  Detent Mec hanism Tolerance Stackup  105
Figure 1.  Schematic of the bearing test facility, not shown to scale.  Load change method The first method to detect the change in end play uses the change in bearing preload from the load cell  shown in Figure 2. The displacement of the outer ring is equal to the measured change in load times the  yield rate of the bearing cartridge system, as shown in equation 4:  /g39PE = /g39T x Y system               Eq (4)  Where /g39T=T 2-T1 is the measured axial thrust load change in lbs, and Y system is the sum of contributions to  the axial yield rate from the two test bearings and the test fixture in /g80in/lb. The yield rate of the bearing  pair was determined using computational bearing analysis tools (DYBA). The bearing pair yield rate takes  into account the change in bearing deflection due to elastic deformation at the Hertzian contact. The yield rate of the system was deter mined through experimental calibration by directly measuring the  displacement for a given force.    The signal from this load cell is collected by a data acquisition system in real time during the bearing test.  To avoid uncertainties in the oute r ring position due to thermal fluctuations, the film thickness  measurements reported here were determined from the nearly instantaneous change in end play when  the bearing rapidly decelerated from full speed to rest. When the full speed and zero speed data points  are collected nearly coincidentally in time, there is no period in which the temperatures can drift to cause  thermal changes in the end play. An example of a test data collection sequence is shown in Figure 3. The data stream nearest to the top of this chart shows ax ial force measured by the load cell. Each of the three  disturbances seen in this data are stop/start cycles during which the bearing deflection was measured.  The full speed thrust load (T 2) was chosen as the data point immediately before stopping the bearing, and  the zero speed load (T 1) was chosen as the data point immediately after the bearing came to rest. During  each restart, we observe  an overshoot transient due to slumping of the lubricant at rest and the  subsequent redistribution after start-up. The difference between these two values is taken as the change  in axial load ( /g39T) to be used in equation 4. One drawback to this technique is that each time the bearing  is stopped, the lubricant distribution is reset. This will tend to prolong the run-in period, and may cause the measured film thicknesses for a given amount of running time to appear thicker than they would if the  bearing had no interruption of use.  110
MPB HSB Spindle Test 60708090100 0 : 0 00 : 0 50 : 1 00 : 1 50 : 2 00 : 2 50 : 3 00 : 3 50 : 4 00 : 4 50 : 5 00 : 5 51 : 0 0 Test Time  (hours:minutes)Temperatures (ºF) 010203040506070 K-RPM; Drag Torque (oz-in)Outer inner shaft Ambient Temp Shaft K-RPM Torque Axial /g205/g3 Fwd./g206 Auto Figure 3.  An example of raw data.  The top data stream shows the axial load that is used to  determine bearing deflections.  The three data streams bundled near the center are, from top to  bottom, temperatures of the outer ring, inner ring, and laboratory ambient.  Air gap displacement change method The total axial deflection of the outer ring is directly measured as the deflection of reference paddles   (/g39D = D 2-D1) connected to the spacers that move with the outer rings, as shown in Figure 2. The axial  position of these paddles is monitored with non-contact proximity (air gap) sensors. To eliminate  uncertainties due to run-out of the bearing system, the response from a pair of paddles mounted on  opposite sides is averaged to produce the total change in axial position. To convert this total change in axial position into a change in the bearing end play fo r use in equation 3, one must first correct for the  displacement due to yield rate of the bearing pair.  To determine the contribution due to the yield rate of the bearing pair, let:  h 1= axial deflection of a single row at thrust load T 1 h2= axial deflection of a single row at thrust load T 2 f1= change in stick-out of a single row due to the EHD film at zero speed = 0  f2= change in stick-out of a single row due to the EHD film at full speed  The axial positions given by the air gap sensor may be written as:  D2 = 2(-h 2 + f 2)              Eq (5a)  D1 = 2(-h 1 + f 1)              Eq (5b)  Or: /g39D = D 2 – D 1 = -2h 2 +2h 1 +2f 2            Eq (6)  So that:  2f2 = /g39D + 2(h 2-h1)             Eq (7)  The quantity 2f 2 is then taken as the change in end play ( /g39PE) to be used in equation 3. The quantity (h 2h1) is equal to the product of the measured change in thrust load and the yield rate of a single bearing. As  mentioned above, this yield rate is determined using computational bearing analysis tools (DYBA). Since  111
this quantity adds to the capacitance measurement distances, the resulting EHD is larger than if this  correction were not used. Results and Discussion  Coray vs. Pennzane Oil When reviewing the results shown belo w, bear in mind that these film thickness results apply only to this  specific set of bearing geometry, materials and operational conditions. Each experiment shown here was  performed with the same pair of 304 sized angular contact bearings operating at 6000 rpm.    The first experiments described here were designed to determine the time scales for run-in, and the final  absolute film thicknesses, of two base oils that have been commonly used in aerospace applications. In  each case, approximately 20 mg of the base oil, formulated with TCP, was added with a clean syringe to  each row. This quantity was sufficien t to show a meniscus at each ball/race interface at 10X. The  bearings were then operated at 6000 rpm. At periodic intervals during the run-in, bearing operation was halted to measure the film thickness. At each interruption, film thickness was recorded and the speed was  returned to 6000 rpm.    Results of these experiments are shown in Figure 4, where the open symbols represent Coray and the  filled symbols represent Pennzane. In both cases, the tr iangles represent data  collected with the load cell  and the circles represent data collected with the displacement sensors. First of all, we find excellent agreement between the two indepe ndent techniques for measuri ng deflection of the rings. Second, we  0 20 40 60 80 100 12005101520253035Total Film Thickness (uin) Time (hrs) Figure 4. Total Film thickness (in microinches) as a function of cumulative time operating at 6000  rpm with neat base oils formulated with TCP.  Open symbols represent Coray and filled symbols  represent Pennzane.  Triangles are the results from the load cell and circles are the results from  the displacement sensors.  find that the run in of Penn zane was nearly complete after only 24 hours, while the Coray took  significantly longer. Third, after 96 hours of cumulative use, the film thickness of Coray was more than  twice that of the Pennzane. More significantly, the total film thickness for Pennzane at the contact  between the ball and inner ring was very small. Using the Hamrock-Dowson expression [7] to determine the partitioning of the film between the inner and outer races, we find that 40% of the film is at the inner  ring and 60% is at the outer. This implies that the inner ring film thickness is approximately 3.6 /g80in after  112
only 24 hours of operation. Since the composite surface roughness for the test articles is approximately  2.0/g80in, the nominal operating condition provides a film that is under 3 times the composite roughness.   The fully flooded total film thickness (the sum of the inner and outer) predicted by the Hamrock-Dowson  expression [7] is 132 /g80in for Pennzane. With the conventional Coy-Zaretsky starvation reduction factors  [8,9], this prediction falls to approximately 77 /g80in for Pennzane. Our measurements show ~9 /g80in total film  thickness after 24 hours. It is evident that our films are much more deeply starved than predicted by the  Coy-Zaretsky analysis. The primary reason for this discrepancy is that the Coy-Zaretsky approach takes  into account kinematic starvation, where resupply of the EHD contact is limited by rapid ball passes, but  neglects deeper starvation due to a restricted amount of lubricant in the vicinity of the contact. In fact, the  experiments used to validate the Coy-Zaretsky  approach were perfo rmed an oil-air mist of lubricant  provided to the contact [9]. Developing the ability to make quantitative estimates of this steady state film  thickness with restricted lubricant supply is the subject of ongoing investigation.  Rheolube Grease Results For the next set of experiments reported here, the bearing was cleaned and a single charge of 1.1 g  Rheolube grease per row (Pennzane oil with a thickener, obtained from Nye Lubricants) was distributed throughout the ball complement. The bearing was then operated at 6000 rpm, and operation was  interrupted sporadically to measure film thickness as described above. The symbols shown in Figure 5  represent the film thickness as measured by capacitance proximity sensors (circles) and load cell (triangles). The first data point was collected after the initial lubricant run-in disturbances were  suppressed after 22 hours. The average value for the total film thickness from the two techniques was 19  /g80in. After this point, the film thickness fell to less than 17 min after 98 hours, but appeared to rise slightly  to 18 /g80in after 365 hours.   0 100 200 300 4000510152025Total Film Thickness (uin) Time (hrs) Figure 5. Total Film thickness (in microinches) as a function of cumulative time operating at 6000  rpm with a grease pack of 1.1 g Rheolube.  Triangles are the results from the load cell and circles  are the results from the displacement sensors.  The observation of a long-term grease film that is much thicker than the oil film contradicts many of the  observations reported in the literature. Others have found that the response with grease is initially much  greater during run-in, but that it falls to the oil film thickness after the run-in is complete and then it falls  113