ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
acceptance. Changing the solenoid design from a single shaft to two solenoids side by side was  considered, but rejected as too drastic a change.  The change selected was to use ball bearings as followers between the solenoid and drive arms. The  space was limited but the small 1.0-mm by 3.2-mm ball bearing could be made to fit. See Figure 9 and  compare to Figure 2. The contact load was high but within the manufacturer's limits of 4,000 MPa (580  ksi). This change was implemented on the subsequent group build, and has been successful with long  ratchet wheel cycles before mechanism failure. There are a few dark particles after a ratchet wheel cycle  test, but they are located at stop pins and not at the bearing followers that actuate the drive arms.  '*  I --3  L 4  1  Figure 9. Ball bearings now transfer the torque from the solenoid to the drive arms  Conclusion  Due to volume constraints and the desire for a simple assembly, single gear teeth were made as integral  part of arms and rotors in a small safety mechanism. The gear teeth were discovered in the prototype to  generate many very small particles that were pulled into the bearings of the adjacent solenoid and caused  very early failures in life tests. The cause of the particle generation was contact stresses exceeding the  yield strength of the gear teeth. The lesson learned was to prioritize our efforts in analysis and failure  mode prediction on new or unfamiliar design concepts that are incorporated into a mechanism. An  ancillary lesson is to consider and evaluate additional concepts when moderate or high risks are identified  in the base design.  35
References  1. D.W. Plummer and W.H. Greenwood (1993). A Primer on Unique Signal Stronglinks, SAND93-0951,  Sandia National Laboratories, Albuquerque, NM.  2. J.E. Shigley and C.R. Mischke (1 989). Mechanical Engineering Design, 5'h edition, McGraw-Hill, New  York City, NY.  36
Radiometer  MCS (Mars Climate Sounder)  This atmospheric profiler will detect vertical variations of temperature, dust, and water vapor  concentrations in the Martian atmosphere.  Radar  SHARAD (Shallow Radar)  This sounding radar will probe beneath the Martian surface to see if water ice is present at depths greater  than one meter.  Enaineerinq  The engineering instruments facilitate spacecraft navigation and communications.  High Gain Antenna  *,  II J VADTR Tnstniment Deck  Once the science phase is completed (two years after the mapping orbit is established), the MRO mission  enters a second phase, communications relay. In this phase, the communication equipment on-board  MRO will be used as a communications relay between the Earth and landed crafts on Mars that may not  have sufficient radio power to communicate directly with Earth on their own. This capability allows landed  crafts to use smaller antennas with reduced mass, improving the lander’s science complement potential.  Due to the mapping nature of the mission, the instrument deck of the spacecraft must always be facing  the surface of Mars. Additional pointing requirements include maintaining sun pointing of the solar panels  and keeping the High Gain Antenna Earth pointed for communication purposes. The solution to this  extreme panel and antenna pointing choreography was to put a two-axis gimbal at each of the  appendages: two solar array wings and one High Gain Antenna.  The path from Earth to Mars orbit and mapping of the surface consists of launch, cruise, orbit insertion,  aerobraking, and mapping phases. The spacecraft configuration of the solar array and the High Gain  Antenna are different for each of these phases. During launch, the solar array wings are folded in half and  the High Gain Antenna is positioned directly over the spacecraft bus to fit into the launch vehicle fairing  (Figure 2). Once MRO is launched and in the cruise phase of the mission, the solar array and High Gain  Antenna are pointing in roughly the same direction to capture sunlight and communicate with Earth  (Figure 3). For Mars orbit Insertion and aerobraking, the appendages are moved slightly from the cruise  configuration to produce an aerodynamically stable configuration. The High Gain Antenna and the solar  array wings are the predominant source of atmospheric drag on the spacecraft and must be positioned to  keep the spacecraft stable throughout the aerobraking maneuvers, which last for about 6 months (see  Figure 4). For the mapping phase of the mission, the solar array wings and High Gain Antenna are almost  38
continuously articulated so the wings remain sun pointed and the High Gain Antenna maintains a lock on  Earth. This continuous motion must be performed while the spacecraft maintains precise pointing for high  resolution imaging and high-speed data transmission to Earth (see Figure 1 for a mapping configuration).  Because of the stringent pointing stability requirements, the gimbals were required to be exceedingly  smooth and quiet.  The different phases for the MRO appendages lead to a large range of requirements for the two-axis  gimbal assemblies. The gimbals must be capable of carrying launch loads through their output bearings  for the launch phase. The cruise phase is very benign with minimal load on the gimbals. The orbit  insertion and aerobraking phases put a significant load on the output bearings and gears, since the loads  have a significant component in the backdriving direction for the gimbal actuators. Once the spacecraft is  in the mapping phase, the high resolution capability of the instruments on board require that the gimbal  assemblies do not produce any significant disturbance to the spacecraft platform while they are  continuously scanning to maintain the required pointing of the attached appendage. The gimbals must  withstand all of these load combinations and still maintain extreme pointing accuracy and smooth  operation once at Mars. In addition to the smooth motion, the lowest structural frequency in the mapping  configuration is determined by the natural frequency of the deployed appendages. A major contributor to  the frequency of the deployed appendage is the gimbal actuator output stiffness for each axis.  Figure 2. MRO Launch  Configuration Figure 3. MRO Cruise  Configuration Figure 4. MRO Orbit  Insertion & Aerobraking  Configuration  Gimbal Actuator Confiauration  Each two-axis gimbal consists of two identical gimbal actuators, structurally connected with application  specific components. The core of the actuator is a 130-mm diameter two-phase brushless dc motor with a  large number of poles in order to maintain smooth rotor velocity. The brushless motor is commutated  using a resolver with the same number of poles as the motor to simplify the commutation logic. The motor  directly drives the wave generator of the output harmonic gear component set through a bellows coupling.  The bellows coupling was used to minimize speed ripple that would cause disturbances while operating.  The harmonic gear is a size-32 HDC, standard-cup-type unit. The flexspline is mounted to the actuator  housing and provides the torque reaction mount. The circular spline is mounted in a pair of preloaded  angular-contact ball bearings. A multi-speed output resolver is installed between the angular-contact  bearings for a compact assembly that measures the output position to the accuracy required for the MRO  mission pointing. This arrangement of drive components provides a zero-backlash actuator with minimal  mechanically generated disturbance sources and applies all of the externally generated loads directly to  the harmonic gear teeth. A photo of a completed flight gimbal actuator assembly is shown in Figure 5.  39
The need for the harmonic gear to take the external loads led to the selection of the size of the harmonic  gear in concert with the required output stiffness. The available volume and mass allocation for the  actuators was minimal on the MRO spacecraft, requiring use of the smallest gears possible to achieve the  required load and stiffness capability. The loads from the Mars orbit insertion and aerobraking phases  needed to be taken into account along with all other applied loads during the mission. The mass of the  solar array wings and the antenna assembly loading the output gear teeth in the acceleration environment  of orbit insertion was one source of applied load. The force from aerodynamic loading on the large area  array and antenna during aerobraking was another source of loading. These conditions together defined  the magnitude of the loads that would be applied to the harmonic gear output teeth in flight. It was  determined that the applied loads could be carried with appropriate margins by a size-32, cup-type  harmonic gear component set.  The output torsional stiffness of the actuator axes affects the spacecraft dynamics. The spacecraft  sensitivity to jitter disturbances and the attitude control system authority dictated a minimum natural  frequency for the deployed solar array wings and High Gain Antenna. Since the gimbal actuator was a  significant contributor to the overall appendage stiffness and the harmonic gear teeth were the load  reacting devices, an output stiffness was required of the chosen harmonic gear that exceeded its  standard specification significantly. The magnitude of this stiff ness increase for the selected size of  harmonic gear was within the range of experience for this type of application on other programs. Since  there was no significant difference in the stiffness requirement when compared to other heritage  programs, this was not considered a significant risk to the program.  The selection of the harmonic gear materials from the available set involved comparing the MRO  requirements with those of previous flight programs in order to maintain as much heritage as possible. A  fundamental tenet for this and many programs was to use only corrosion resistant materials in all space  mechanisms as is commonly done in the medical, semiconductor, and food processing industries. The  materials of the various pieces of the MRO harmonic gear component set are listed in Table 1. This  combination of materials had been used successfully in several programs with stiff ness and total lifetime  revolutions requirements that were similar to MRO. Table 2 lists the other heritage applications of the  same material combination with similar functional requirements. To maximize the life cycle capability of all  40
Figure 6. Engineering Development Unit Configuration in Vacuum Chamber  Flexsp I i n e Table 2. Materials and Surface Finishes of Heritage Hardware from Previous  Programs with Similar Stiffness and Life requirements as the MRO Applications  Lubricant Program  (Harmonic Drive Size)  MRO  (size 32) 15-5 PH H1150  Melonite  15-5 PH H1150  Melonite  15-5 PH H1075  Melonite  15-5PH H1075 (size 25) Penzane  2001 -3Pb Rheolube  2004  Braycote 602  Bray 81 52 oil  Penzane  2001 -3Pb Rheolube  2004  Penzane  2001 -3Pb Rheolube Circular Spline  Program #5  (size 32) 15-5PH H1075  2004  15-5 PH H1075 15-5 PH H1150 Rheolube 2000 with  Melonite 3% lead Napthenate 15-5 PH H1075  15-5PH H1150  15-5 PH H1150  Melonite  (sbe 40) I Melonite I I 2001 -3Pb Rheolube  42
The EDU actuator operated up to 6.1 million input revolutions, when the output telemetry indicated the  actuator was not following the input signals properly. Later analysis revealed there were indications of  improper operation as early as 4 million input revolutions that were not diagnosed due to limitations in the  test setup with the stepper motor driver. The EDU actuator was removed from the chamber,  disassembled, and inspected. Figures 7, 8, and 9 show the condition of the harmonic gear teeth at this  inspection. Note that the tooth profile was completely obliterated and damaged across the entire width of  the flexspline teeth and most of the width of the circular spline.  I  i Intacl  Figure 7. EDU Circular Spline Life Damage Figure 8. EDU Flexspline Life Damage  Figure 9. Magnification of EDU Flexspline Life Test Damage  43
Once the dramatic failure of the harmonic gear assembly was seen, a detailed review of the  manufacturer’s documents uncovered that there were some problems with the Melonite coating on the  group of flexsplines that included the life unit, labeled Lot B. Two additional Melonite processing groups of  flexsplines had been received at the time of the failure, labeled Lot C and Lot D. Operation of these units  was started to determine if the Melonite processing on the EDU life test unit was the source of the failure.  A harmonic assembly from each of the two remaining Melonite process groups was placed in a harmonic  gear fixture (not in the gimbal actuator) and operated in vacuum with thermal cycling over a 12-hour  period, an output load, and an input speed of 130 RPM. Additionally, one of the two units was tested  using the same Penzane family of lubricants as the EDU and the other was tested using Braycote@ 602.  Both assemblies failed at approximately two million input revolutions. Figure 10 shows the failed splines  that were tested with the Penzane family of lubricants. Figure 11 shows the failed splines that were tested  with the Braycote@ 602 lubricant. A significant result of these two tests was that the lubricant type made  no difference at all, with both assemblies failing at nearly the same number of revolutions with the same  type and level of damage.  I  Figure 10. Circular Spline and Flexspline from Harmonic Only Test #1  Tested with Penzane  Figure 11. Circular Spline and Flexspline from Harmonic Only Test #I2  Tested with Braycote@ 602  44
Since the lubricant type made absolutely no difference in the life capability of the harmonic drive gear  teeth, it was clear the problem was related directly to the material combination of the components and  their internal stress level. The appearance of the failure surfaces gave the impression of a possible galling  condition at work, but it was not clear if galling was the initiator of the failure or a consequence of the  damage once the failure had been initiated. Rough mathematical analyses were performed to estimate  the contact stress at the harmonic gear teeth from the preload and under the applied load in the  operation. The constant external load in the harmonic fixture runs represented the worst-case load from  the cable management system and was responsible for a roughly 40% increase in tooth contact stresses  over the preload. The estimated contact stresses in the EDU life test were in the realm of 750 MPa. The  galling threshold listed for a 15-5 PH stainless steel contact pair is around 14 MPa. This indicated that the  contact stresses compared to the galling threshold for the selected materials was a strong candidate for  the cause of the failures. To minimize the schedule time to a solution, the next group of tests used  Nitronic 60 (another available harmonic gear material) with a listed galling threshold value greater than  345 MPa. Nitronic 60 was identified as a candidate in addition to other standard commercial materials for  the harmonic gear assemblies, like nodular iron. Also, the output stiffness of the assembly was reduced  so the internal harmonic gear preload (and with it the internal tooth stresses) could be reduced as well. At  this point in the project schedule, the flight solar array panels had been fabricated and their stiffness was  measured, allowing reduction of the stiffness margin for these panels in the MRO spacecraft stability  analysis. This made it possible to reduce the gimbal actuator stiffness requirement significantly.  To investigate these issues within the remaining program schedule, three readily available harmonic drive  assemblies were procured with different material combinations and tested on the harmonic gear fixture.  The material combinations consisted of a unit with a nodular iron circular spline and an E4340 flexspline.  The second unit had a Nitronic 60 circular spline and an E4340 flexspline. The third assembly was  composed of a Nitronic 60 circular spline with an E4340 flexspline processed with a Melonite surface.  Additionally, the internal preload of the harmonic gear (to obtain the new required output stiffness) was  reduced in order to lower the gear tooth internal stresses. Table 3 lists all of the life test units and the  material and lubrication configuration for each.  F Table 3. Selected Harmonic Gear Material and Lubricant Configurations  te: LM stands for Lockheed Martin  45
The test units for LM Harmonic Assemblies #3, #4, and #5 were tested in the harmonic drive test fixture,  in the environment, with the applied load and input speed listed in Table 4. The stiffness of the units was  measured at the start of the testing as shown in column 2 of Table 4. The life tests already discussed are  listed in Table 4 for completeness.  Test Unit Table 4. Life Unit Stiff ness, Environments, Conditions, and Revolutions at Failure  Initial Stiffness  with 45/140 N.m Applied Approx.  Applied Torque Load Input Revolutions  (N-mlrad) Environment on Output Speed to failure  vendor  #I (in actuator)  LM Harmonic  Assembly #1  (same unit as  Harmonic drive  vendor Unit #1)  LM Harmonic  Assembly #2  LM Harmonic  Assembly #3  LM Harmonic  Assembly #4 Approx. 68,000  > 56,000  Approx. 68,000  69,000/75,000  56,000  55,600/60300  56,500 (28 N.m)  43,800 (28 N.m)  54,600/61,000  43,800 (28 N.m)  37,300 to 46,300  (28 N.m) Ambient  Vacuum, -25"C/+4OoC  18 hour cycles  Vacuum, -1 5"C/+40°C  12 hour cycles  Vacuum, -1 5"C/+40°C  12 hour cycles  Vacuum, -1 5"C/+4OoC  12 hour cycles  First 3M revs  Vacuum, -1 5"C/+40°C  12 & 24 hour cycles  Next 3M revs  Vacuum, +23"C  constant  Next 8M revs  Vacuum, -1 0°C  constant  Vacuum, -lO"C/O"C  constant  Vacuum, -1 O"C/+l 0°C  constant 11.3 N.m  HGA Cable  Harness  11.3 N.m  11.3 Nom  11.3 N.m  11.3 N.m  11 -3 N.m  HGA Cable  Harness 1750 RPM  125 RPM  & 65 RPM  130 RPM  140 RPM  140 RPM  140 RPM  140 RPM  60 RPM No Failure  4M  2M  2M  4M  No Failure Q  14M  No Failure Q  9.1 M  No Failure Q  13M  Key results of the above are as follows:  0 The high input speed of the harmonic gear only operation performed at the harmonic drive vendor  allows the lubricant to support high contact stresses that would otherwise result in complete  failure of the gear teeth.  The difference from the EDU Assembly #1 unit and the LM Harmonic Assembly #l was the  Melonite lot and coating details. The results of the LM Harmonic Assembly Unit #1 showed that  the Melonite coating was not involved in the failure.  The applied output load, while small when compared to the maximum torque capability of the  harmonic gear, made a difference in the revolution life to failure. Note that the EDU life unit  exhibited failure around 4 million revolutions, while assemblies #1 and #2, with a constant applied  load, failed at 2 million revolutions. This difference could be due to the EDU assembly #1 using a  cable wrap harness, with a variable load depending on output position, compared to a constant  load. 0  0  46
LM Harmonic Assemblies #1 and #2, with failure at a similar number of revolutions using very  different lubricants, indicated the lubricant was not a significant player in the failure mechanism.  This eliminated the lubricant as a variable in further failure investigation.  The LM Harmonic Assembly #3, with nodular iron and a reduced preload, had a longer life to  failure than the stainless steel, supporting the galling hypothesis. Note the regions of damage in  Figure 12.  The LM Harmonic Assemblies #4 and #5 used Nitronic 60 for the circular spline, the highest  galling threshold material that could be obtained in a harmonic gear assembly. The late date of  the testing permitted a reduction of the output stiffness to two-thirds of the initial values used for  the gimbal actuators. The stiffness of the unit in Harmonic Assembly #4 was reduced as the  operation progressed and a method of setting the harmonic gear preload was established. The  result of the run was no failure at all with some minor wear of the harmonic gear teeth, as shown  in Figure 13. The schedule dictated that the first successful combination be used, so the final  material selection was a Nitronic 60 circular spline and a flexspline of E4340 with no additional  surface processing. Note that schedule dictated changing more than one variable at a time.  The setting of the internal preload of the harmonic gear assembly during assembly at the  harmonic drive vendor was critical in achieving the required output stiff ness without compromising  the reliability of the harmonic gear assembly. 0  0  Figure 12. Circular Spline and Flexspline fromHarmonic Only Tex#3  .A n.ri+h +ha ,-ir,.nalmr rnlins A  =:, ('.  Figure 13. Circular Spline and Flexspline trom harmonic vnty I esr lf4  Note the slight wear region on the circular spline showing where the flexspline was engaged with it.  47
Harmonic Gear Assemblv Internal Preload Settinq  The cup type harmonic gear assembly has zero backlash due to the angle of approach, along the rotation  axis, of the flexspline teeth relative to the circular spline teeth (Figure 14). This arrangement leads to a  two-sloped stiffness curve of the output of a harmonic gear. As torque is applied to the gear, the cup  flexes to allow more of the width of the teeth to engage with the circular spline. As more tooth area is  engaged, more load sharing occurs and the stiffness increases. Once the angle between the teeth of the  splines is reduced to near zero, the stiffness has reached its maximum value. As the torque is increased  from this point, the stiffness is related to deflection of the individual teeth and the change in roundness of  the housing and circular spline. Figure 15 shows a stiffness curve for a typical harmonic gear assembly  prior to increasing the internal preload for stiff ness improvement.  Flexspline  tooth angle - Cup flexes to allow more  tooth face engagement / /Flexspline  Circular Spline  tooth angle  Figure 14. Harmonic Drive Cross Section Showing Significant Source of Variable Stiffness  Note the largest stresses on flexspline teeth occur at the point of initial contact, the open end of the cup.  Drawing courtesy of Harmonic Drive, LLC  Increasing the output stiffness of a harmonic drive assembly involves increasing the diameter of the wave  generator in the area where it forces contact between the teeth of the two splines. This is accomplished  by using a different wave generator plug with a larger major diameter of the oval. As the diameter is  increased, the flexspline cup is deflected in the direction of engaging more of the face of the teeth. This  has the same effect as increasing the torque on a nominal unit in the low stiffness region. As the wave  generator plug size is increased, the low stiffness region gets smaller. This trend continues until the  stiffness curve is essentially straight. At this point, a further increase in the diameter of the wave  generator plug will increase the overall gear assembly stiffness and significantly increase the internal  tooth stresses. Figures 16 through 19 show the how the shape of the stiffness curve changes with  different wave generator plugs in a harmonic drive assembly. The low stiffness region is very evident in  Figure 16. A larger plug, after insertion into the wave generator bearing may nearly eliminate the low  stiffness region, as seen in Figure 17. Figures 18 and 19 show the next two larger size wave generator  plugs, without a significant change in the shape of the stiffness curve. The desired operating point for the  flight plug is smallest wave generator plug that exhibits a fairly straight curve. If there is any uncertainty  between units, the smaller one would always be installed to guarantee that the gear teeth were not being  jammed together with high, and unknown, internal stresses. The wave generator plug used in this  example would be Figure 17’s.  48